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Fluid Mechanics: Fundamentals and Applications
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cen72367_fm.qxd 11/23/04 11:22 AM Page i
FLUID MECHANICS FUNDAMENTALS AND APPLICATIONS
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McGRAW-HILL SERIES IN MECHANICAL ENGINEERING Alciatore and Histand: Anderson: Anderson: Anderson: Anderson: Barber: Beer/Johnston: Beer/Johnston/DeWolf: Borman and Ragland: Budynas: Çengel and Boles: Çengel and Cimbala: Çengel and Turner: Çengel: Crespo da Silva: Dieter: Dieter: Doebelin: Dunn: EDS, Inc.: Hamrock/Jacobson/Schmid: Henkel and Pense: Heywood: Holman: Holman: Hsu: Hutton: Kays/Crawford/Weigand: Kelly: Kreider/Rabl/Curtiss: Mattingly: Meirovitch: Norton: Palm: Reddy: Ribando: Schaffer et al.: Schey: Schlichting: Shames: Shigley/Mischke/Budynas: Smith: Stoecker: Suryanarayana and Arici: Turns: Ugural: Ugural: Ullman: Wark and Richards: White: White: Zeid:
Introduction to Mechatronics and Measurement Systems Computational Fluid Dynamics: The Basics with Applications Fundamentals of Aerodynamics Introduction to Flight Modern Compressible Flow Intermediate Mechanics of Materials Vector Mechanics for Engineers Mechanics of Materials Combustion Engineering Advanced Strength and Applied Stress Analysis Thermodynamics: An Engineering Approach Fluid Mechanics: Fundamentals and Applications Fundamentals of Thermal-Fluid Sciences Heat Transfer: A Practical Approach Intermediate Dynamics Engineering Design: A Materials & Processing Approach Mechanical Metallurgy Measurement Systems: Application & Design Measurement & Data Analysis for Engineering & Science I-DEAS Student Guide Fundamentals of Machine Elements Structure and Properties of Engineering Material Internal Combustion Engine Fundamentals Experimental Methods for Engineers Heat Transfer MEMS & Microsystems: Manufacture & Design Fundamentals of Finite Element Analysis Convective Heat and Mass Transfer Fundamentals of Mechanical Vibrations The Heating and Cooling of Buildings Elements of Gas Turbine Propulsion Fundamentals of Vibrations Design of Machinery System Dynamics An Introduction to Finite Element Method Heat Transfer Tools The Science and Design of Engineering Materials Introduction to Manufacturing Processes Boundary-Layer Theory Mechanics of Fluids Mechanical Engineering Design Foundations of Materials Science and Engineering Design of Thermal Systems Design and Simulation of Thermal Systems An Introduction to Combustion: Concepts and Applications Stresses in Plates and Shells Mechanical Design: An Integrated Approach The Mechanical Design Process Thermodynamics Fluid Mechanics Viscous Fluid Flow Mastering CAD/CAM
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FLUID MECHANICS FUNDAMENTALS AND APPLICATIONS
YUNUS A. ÇENGEL Department of Mechanical Engineering University of Nevada, Reno
JOHN M. CIMBALA Department of Mechanical and Nuclear Engineering The Pennsylvania State University
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FLUID MECHANICS: FUNDAMENTALS AND APPLICATIONS Published by McGraw-Hill, a business unit of The McGraw-Hill Companies, Inc., 1221 Avenue of the Americas, New York, NY 10020. Copyright © 2006 by The McGraw-Hill Companies, Inc. All rights reserved. No part of this publication may be reproduced or distributed in any form or by any means, or stored in a database or retrieval system, without the prior written consent of The McGraw-Hill Companies, Inc., including, but not limited to, in any network or other electronic storage or transmission, or broadcast for distance learning. Some ancillaries, including electronic and print components, may not be available to customers outside the United States. This book is printed on acid-free paper. 1 2 3 4 5 6 7 8 9 0 DOW/DOW 0 9 8 7 6 5 4 ISBN 0–07–247236–7 Senior Sponsoring Editor: Suzanne Jeans Managing Developmental Editor: Debra D. Matteson Developmental Editor: Kate Scheinman Senior Marketing Manager: Mary K. Kittell Senior Project Manager: Sheila M. Frank Senior Production Supervisor: Sherry L. Kane Media Technology Producer: Eric A. Weber Senior Designer: David W. Hash (USE) Cover image: © Getty/Eric Meola, Niagara Falls Senior Photo Research Coordinator: Lori Hancock Photo Research: Judy Ladendorf/The Permissions Group Supplemental Producer: Brenda A. Ernzen Compositor: Lachina Publishing Services Typeface: 10.5/12 Times Roman Printer: R. R. Donnelley Willard, OH
Library of Congress Cataloging-in-Publication Data Çengel, Yunus A.
Fluid mechanics : fundamentals and applications / Yunus A. Çengel, John M. Cimbala.—1st ed. p. cm.—(McGraw-Hill series in mechanical engineering) ISBN 0–07–247236–7 1. Fluid dynamics. I. Cimbala, John M. II. Title. III. Series. TA357.C43 2006 620.1'06—dc22
www.mhhe.com
2004058767 CIP
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Dedication To all students—In hopes of enhancing your desire and enthusiasm to explore the inner workings of our marvelous universe, of which fluid mechanics is a small but fascinating part; our hope is that this book enhances your love of learning, not only about fluid mechanics, but about life.
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ABOUT
THE
AUTHORS
Yunus A. Çengel
is Professor Emeritus of Mechanical Engineering at the University of Nevada, Reno. He received his B.S. in mechanical engineering from Istanbul Technical University and his M.S. and Ph.D. in mechanical engineering from North Carolina State University. His research areas are renewable energy, desalination, exergy analysis, heat transfer enhancement, radiation heat transfer, and energy conservation. He served as the director of the Industrial Assessment Center (IAC) at the University of Nevada, Reno, from 1996 to 2000. He has led teams of engineering students to numerous manufacturing facilities in Northern Nevada and California to do industrial assessments, and has prepared energy conservation, waste minimization, and productivity enhancement reports for them. Dr. Çengel is the coauthor of the widely adopted textbook Thermodynamics: An Engineering Approach, 4th edition (2002), published by McGraw-Hill. He is also the author of the textbook Heat Transfer: A Practical Approach, 2nd edition (2003), and the coauthor of the textbook Fundamentals of ThermalFluid Sciences, 2nd edition (2005), both published by McGraw-Hill. Some of his textbooks have been translated to Chinese, Japanese, Korean, Spanish, Turkish, Italian, and Greek. Dr. Çengel is the recipient of several outstanding teacher awards, and he has received the ASEE Meriam/Wiley Distinguished Author Award for excellence in authorship in 1992 and again in 2000. Dr. Çengel is a registered Professional Engineer in the State of Nevada, and is a member of the American Society of Mechanical Engineers (ASME) and the American Society for Engineering Education (ASEE).
John M. Cimbala is Professor of Mechanical Engineering at The Pennsylvania State Univesity, University Park. He received his B.S. in Aerospace Engineering from Penn State and his M.S. in Aeronautics from the California Institute of Technology (CalTech). He received his Ph.D. in Aeronautics from CalTech in 1984 under the supervision of Professor Anatol Roshko, to whom he will be forever grateful. His research areas include experimental and computational fluid mechanics and heat transfer, turbulence, turbulence modeling, turbomachinery, indoor air quality, and air pollution control. During the academic year 1993–94, Professor Cimbala took a sabbatical leave from the University and worked at NASA Langley Research Center, where he advanced his knowledge of computational fluid dynamics (CFD) and turbulence modeling. Dr. Cimbala is the coauthor of the textbook Indoor Air Quality Engineering: Environmental Health and Control of Indoor Pollutants (2003), published by Marcel-Dekker, Inc. He has also contributed to parts of other books, and is the author or co-author of dozens of journal and conference papers. More information can be found at www.mne.psu.edu/cimbala. Professor Cimbala is the recipient of several outstanding teaching awards and views his book writing as an extension of his love of teaching. He is a member of the American Institute of Aeronautics and Astronautics (AIAA), the American Society of Mechanical Engineers (ASME), the American Society for Engineering Education (ASEE), and the American Physical Society (APS).
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BRIEF CONTENTS CHAPTER
ONE
INTRODUCTION AND BASIC CONCEPTS
1
CHAPTER TWO PROPERTIES OF FLUIDS
35
CHAPTER THREE PRESSURE AND FLUID STATICS
CHAPTER
FOUR
FLUID KINEMATICS
CHAPTER
65
121
FIVE
MASS, BERNOULLI, AND ENERGY EQUATIONS
CHAPTER
SIX
MOMENTUM ANALYSIS OF FLOW SYSTEMS
CHAPTER
171 227
SEVEN
DIMENSIONAL ANALYSIS AND MODELING
CHAPTER
EIGHT
FLOW IN PIPES
321
CHAPTER
NINE
DIFFERENTIAL ANALYSIS OF FLUID FLOW
269
399
CHAPTER TEN APPROXIMATE SOLUTIONS OF THE NAVIER–STOKES EQUATION
CHAPTER
ELEVEN
FLOW OVER BODIES: DRAG AND LIFT
561
C H A P T E R T W E LV E COMPRESSIBLE FLOW
611
CHAPTER THIRTEEN OPEN-CHANNEL FLOW
CHAPTER
FOURTEEN
TURBOMACHINERY
CHAPTER
679
735
FIFTEEN
INTRODUCTION TO COMPUTATIONAL FLUID DYNAMICS
817
471
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CONTENTS Preface
xv
CHAPTER
Application Spotlight: What Nuclear Blasts and Raindrops Have in Common 31
INTRODUCTION AND BASIC CONCEPTS 1–1
1
Introduction 2
CHAPTER
What Is a Fluid? 2 Application Areas of Fluid Mechanics
1–2 1–3 1–4
The No-Slip Condition 6 A Brief History of Fluid Mechanics 7 Classification of Fluid Flows 9
2–1
2–6 2–7
1–8
44
Viscosity 46 Surface Tension and Capillary Effect 51 53
Summary 55 References and Suggested Reading
56
Application Spotlight: Cavitation 57 Problems
21
58
Problem-Solving Technique 22 Step 1: Problem Statement 22 Step 2: Schematic 23 Step 3: Assumptions and Approximations 23 Step 4: Physical Laws 23 Step 5: Properties 23 Step 6: Calculations 23 Step 7: Reasoning, Verification, and Discussion
1–9
38
Vapor Pressure and Cavitation 39 Energy and Specific Heats 41 Coefficient of Compressibility 42
Capillary Effect
Mathematical Modeling of Engineering Problems 21 Modeling in Engineering
Density and Specific Gravity 37
Coefficient of Volume Expansion
System and Control Volume 14 Importance of Dimensions and Units 15 Some SI and English Units 16 Dimensional Homogeneity 18 Unity Conversion Ratios 20
1–7
36
Density of Ideal Gases
2–3 2–4 2–5
35
Introduction 36 Continuum
2–2
30
TWO
PROPERTIES OF FLUIDS
4
Viscous versus Inviscid Regions of Flow 9 Internal versus External Flow 10 Compressible versus Incompressible Flow 10 Laminar versus Turbulent Flow 11 Natural (or Unforced) versus Forced Flow 11 Steady versus Unsteady Flow 11 One-, Two-, and Three-Dimensional Flows 12
1–5 1–6
Summary 30 References and Suggested Reading Problems 32
ONE
CHAPTER
PRESSURE AND FLUID STATICS 3–1
25
1–10 Accuracy, Precision, and Significant Digits 26
3–2
68
The Manometer 71 Other Pressure Measurement Devices
3–3 3–4
65
Pressure 66 Pressure at a Point 67 Variation of Pressure with Depth
23
Engineering Software Packages 24 Engineering Equation Solver (EES) FLUENT 26
THREE
74
The Barometer and Atmospheric Pressure 75 Introduction to Fluid Statics 78
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ix CONTENTS
3–5
Hydrostatic Forces on Submerged Plane Surfaces 79 Special Case: Submerged Rectangular Plate
3–6 3–7
Summary 102 References and Suggested Reading Problems 103
CHAPTER FLUID KINEMATICS 4–1
5–1
5–2 97
103
5–3 5–4
FOUR
Lagrangian and Eulerian Descriptions 122
Fundamentals of Flow Visualization 129
5–5 5–6
5–7
4–5
Summary 215 References and Suggested Reading Problems 216
Other Kinematic Descriptions 139
CHAPTER
158
SIX
MOMENTUM ANALYSIS OF FLOW SYSTEMS 227 155
6–1
Application Spotlight: Fluidic Actuators 157 Summary 156 References and Suggested Reading Problems 158
216
139
The Reynolds Transport Theorem 148 Alternate Derivation of the Reynolds Transport Theorem 153 Relationship between Material Derivative and RTT
Energy Analysis of Steady Flows 206 Special Case: Incompressible Flow with No Mechanical Work Devices and Negligible Friction 208 Kinetic Energy Correction Factor, a 208
Plots of Fluid Flow Data 136
Types of Motion or Deformation of Fluid Elements Vorticity and Rotationality 144 Comparison of Two Circular Flows 147
Applications of the Bernoulli Equation 194 General Energy Equation 201 Energy Transfer by Heat, Q 202 Energy Transfer by Work, W 202
Profile Plots 137 Vector Plots 137 Contour Plots 138
4–4
Mechanical Energy and Efficiency 180 The Bernoulli Equation 185 Acceleration of a Fluid Particle 186 Derivation of the Bernoulli Equation 186 Force Balance across Streamlines 188 Unsteady, Compressible Flow 189 Static, Dynamic, and Stagnation Pressures 189 Limitations on the Use of the Bernoulli Equation 190 Hydraulic Grade Line (HGL) and Energy Grade Line (EGL) 192
121
Streamlines and Streamtubes 129 Pathlines 130 Streaklines 132 Timelines 134 Refractive Flow Visualization Techniques 135 Surface Flow Visualization Techniques 136
4–3
Conservation of Mass 173 Mass and Volume Flow Rates 173 Conservation of Mass Principle 175 Moving or Deforming Control Volumes 177 Mass Balance for Steady-Flow Processes 177 Special Case: Incompressible Flow 178
Acceleration Field 124 Material Derivative 127
4–2
Introduction 172 Conservation of Mass 172 Conservation of Momentum 172 Conservation of Energy 172
92
Fluids in Rigid-Body Motion 95 Special Case 1: Fluids at Rest 96 Special Case 2: Free Fall of a Fluid Body Acceleration on a Straight Path 97 Rotation in a Cylindrical Container 99
FIVE
MASS, BERNOULLI, AND ENERGY EQUATIONS 171
82
Hydrostatic Forces on Submerged Curved Surfaces 85 Buoyancy and Stability 89 Stability of Immersed and Floating Bodies
3–8
CHAPTER
6–2 6–3
Newton’s Laws and Conservation of Momentum 228 Choosing a Control Volume 229 Forces Acting on a Control Volume 230
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x FLUID MECHANICS
6–4
8–3
The Linear Momentum Equation 233 Special Cases 235 Momentum-Flux Correction Factor, b 235 Steady Flow 238 Steady Flow with One Inlet and One Outlet 238 Flow with No External Forces 238
6–5 6–6
Entry Lengths
8–4
8–5
8–6 8–7
259
SEVEN
DIMENSIONAL ANALYSIS AND MODELING 7–1 7–2 7–3 7–4
7–5
269
272
Dimensional Analysis and Similarity 277 The Method of Repeating Variables and the Buckingham Pi Theorem 281 Historical Spotlight: Persons Honored by Nondimensional Parameters 289 Experimental Testing and Incomplete Similarity 297 Setup of an Experiment and Correlation of Experimental Data 297 Incomplete Similarity 298 Wind Tunnel Testing 298 Flows with Free Surfaces 301
Application Spotlight: How a Fly Flies 304 Summary 305 References and Suggested Reading Problems 305
CHAPTER FLOW IN PIPES 8–1 8–2
305
324
Flow Rate and Velocity Measurement 364
Summary 384 References and Suggested Reading Problems 386
CHAPTER
385
NINE
DIFFERENTIAL ANALYSIS OF FLUID FLOW 9–1 9–2
399
Introduction 400 Conservation of Mass—The Continuity Equation 400 Derivation Using the Divergence Theorem 401 Derivation Using an Infinitesimal Control Volume 402 Alternative Form of the Continuity Equation 405 Continuity Equation in Cylindrical Coordinates 406 Special Cases of the Continuity Equation 406
EIGHT
Introduction 322 Laminar and Turbulent Flows 323
356
Application Spotlight: How Orifice Plate Flowmeters Work, or Do Not Work 383
321
Reynolds Number
Minor Losses 347 Piping Networks and Pump Selection 354
Pitot and Pitot-Static Probes 365 Obstruction Flowmeters: Orifice, Venturi, and Nozzle Meters 366 Positive Displacement Flowmeters 369 Turbine Flowmeters 370 Variable-Area Flowmeters (Rotameters) 372 Ultrasonic Flowmeters 373 Electromagnetic Flowmeters 375 Vortex Flowmeters 376 Thermal (Hot-Wire and Hot-Film) Anemometers 377 Laser Doppler Velocimetry 378 Particle Image Velocimetry 380
Dimensions and Units 270 Dimensional Homogeneity 271 Nondimensionalization of Equations
Turbulent Flow in Pipes 335
Piping Systems with Pumps and Turbines
8–8
CHAPTER
Laminar Flow in Pipes 327
Turbulent Shear Stress 336 Turbulent Velocity Profile 338 The Moody Chart 340 Types of Fluid Flow Problems 343
253
Summary 259 References and Suggested Reading Problems 260
326
Pressure Drop and Head Loss 329 Inclined Pipes 331 Laminar Flow in Noncircular Pipes 332
Review of Rotational Motion and Angular Momentum 248 The Angular Momentum Equation 250 Special Cases 252 Flow with No External Moments Radial-Flow Devices 254
The Entrance Region 325
9–3
The Stream Function 412 The Stream Function in Cartesian Coordinates 412 The Stream Function in Cylindrical Coordinates 419 The Compressible Stream Function 420
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xi CONTENTS
9–4
Conservation of Linear Momentum—Cauchy’s Equation 421 Derivation Using the Divergence Theorem 421 Derivation Using an Infinitesimal Control Volume Alternative Form of Cauchy’s Equation 425 Derivation Using Newton’s Second Law 425
9–5
422
The Navier–Stokes Equation 426 Introduction 426 Newtonian versus Non-Newtonian Fluids 427 Derivation of the Navier–Stokes Equation for Incompressible, Isothermal Flow 428 Continuity and Navier–Stokes Equations in Cartesian Coordinates 430 Continuity and Navier–Stokes Equations in Cylindrical Coordinates 431
9–6
Differential Analysis of Fluid Flow Problems 432 Calculation of the Pressure Field for a Known Velocity Field 432 Exact Solutions of the Continuity and Navier–Stokes Equations 437 Summary 455 References and Suggested Reading Problems 456
456
10–6 The Boundary Layer Approximation 510 The Boundary Layer Equations 515 The Boundary Layer Procedure 520 Displacement Thickness 524 Momentum Thickness 527 Turbulent Flat Plate Boundary Layer 528 Boundary Layers with Pressure Gradients 534 The Momentum Integral Technique for Boundary Layers 539
Application Spotlight: Droplet Formation 549 Summary 547 References and Suggested Reading Problems 550
CHAPTER
548
ELEVEN
FLOW OVER BODIES: DRAG AND LIFT
561
11–1 Introduction 562 11–2 Drag and Lift 563 11–3 Friction and Pressure Drag 567 Reducing Drag by Streamlining Flow Separation 569
568
11–4 Drag Coefficients of Common Geometries 571
CHAPTER
TEN
APPROXIMATE SOLUTIONS OF THE NAVIER–STOKES EQUATION 471
Biological Systems and Drag Drag Coefficients of Vehicles Superposition 577
11–5 Parallel Flow over Flat Plates 579 Friction Coefficient
10–1 Introduction 472 10–2 Nondimensionalized Equations of Motion 473 10–3 The Creeping Flow Approximation 476 Drag on a Sphere in Creeping Flow
479
10–4 Approximation for Inviscid Regions of Flow 481 Derivation of the Bernoulli Equation in Inviscid Regions of Flow 482
572 574
580
11–6 Flow over Cylinders and Spheres 583 Effect of Surface Roughness
586
11–7 Lift 587 End Effects of Wing Tips 591 Lift Generated by Spinning 594
Application Spotlight: Drag Reduction 600 Summary 598 References and Suggested Reading Problems 601
599
10–5 The Irrotational Flow Approximation 485 Continuity Equation 485 Momentum Equation 487 Derivation of the Bernoulli Equation in Irrotational Regions of Flow 487 Two-Dimensional Irrotational Regions of Flow 490 Superposition in Irrotational Regions of Flow 494 Elementary Planar Irrotational Flows 494 Irrotational Flows Formed by Superposition 501
CHAPTER
T W E LV E
COMPRESSIBLE FLOW
611
12–1 Stagnation Properties 612 12–2 Speed of Sound and Mach Number 615 12–3 One-Dimensional Isentropic Flow 617 Variation of Fluid Velocity with Flow Area 620 Property Relations for Isentropic Flow of Ideal Gases
622
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xii FLUID MECHANICS Summary 723 References and Suggested Reading Problems 725
12–4 Isentropic Flow through Nozzles 624 Converging Nozzles 625 Converging–Diverging Nozzles
629
724
12–5 Shock Waves and Expansion Waves 633 Normal Shocks 633 Oblique Shocks 640 Prandtl–Meyer Expansion Waves
CHAPTER TURBOMACHINERY
644
12–6 Duct Flow with Heat Transfer and Negligible Friction (Rayleigh Flow) 648 Property Relations for Rayleigh Flow Choked Rayleigh Flow 655
654
660
Application Spotlight: Shock-Wave/ Boundary-Layer Interactions 667 Summary 668 References and Suggested Reading Problems 669
735
14–1 Classifications and Terminology 736 14–2 Pumps 738 Pump Performance Curves and Matching a Pump to a Piping System 739 Pump Cavitation and Net Positive Suction Head 745 Pumps in Series and Parallel 748 Positive-Displacement Pumps 751 Dynamic Pumps 754 Centrifugal Pumps 754 Axial Pumps 764
12–7 Adiabatic Duct Flow with Friction (Fanno Flow) 657 Property Relations for Fanno Flow Choked Fanno Flow 663
FOURTEEN
14–3 Pump Scaling Laws 773 Dimensional Analysis 773 Pump Specific Speed 775 Affinity Laws 777
669
14–4 Turbines 781
CHAPTER THIRTEEN OPEN-CHANNEL FLOW
Positive-Displacement Turbines Dynamic Turbines 782 Impulse Turbines 783 Reaction Turbines 785
679
13–1 Classification of Open-Channel Flows 680 Uniform and Varied Flows 680 Laminar and Turbulent Flows in Channels
14–5 Turbine Scaling Laws 795 Dimensionless Turbine Parameters Turbine Specific Speed 797 Gas and Steam Turbines 800
681
13–2 Froude Number and Wave Speed 683 Speed of Surface Waves
782
Application Spotlight: Rotary Fuel Atomizers 802
685
13–3 Specific Energy 687 13–4 Continuity and Energy Equations 690 13–5 Uniform Flow in Channels 691 Critical Uniform Flow 693 Superposition Method for Nonuniform Perimeters
795
Summary 803 References and Suggested Reading Problems 804 693
13–6 Best Hydraulic Cross Sections 697
CHAPTER
803
FIFTEEN
INTRODUCTION TO COMPUTATIONAL FLUID DYNAMICS 817
Rectangular Channels 699 Trapezoidal Channels 699
13–7 Gradually Varied Flow 701 Liquid Surface Profiles in Open Channels, y (x) Some Representative Surface Profiles 706 Numerical Solution of Surface Profile 708
703
13–8 Rapidly Varied Flow and Hydraulic Jump 709 13–9 Flow Control and Measurement 714 Underflow Gates 714 Overflow Gates 716
15–1 Introduction and Fundamentals 818 Motivation 818 Equations of Motion 818 Solution Procedure 819 Additional Equations of Motion 821 Grid Generation and Grid Independence Boundary Conditions 826 Practice Makes Perfect 830
821
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xiii CONTENTS
15–2 Laminar CFD Calculations 831
TABLE A–11
Pipe Flow Entrance Region at Re 500 831 Flow around a Circular Cylinder at Re 150 833
15–3 Turbulent CFD Calculations 840 Flow around a Circular Cylinder at Re 10,000 843 Flow around a Circular Cylinder at Re 107 844 Design of the Stator for a Vane-Axial Flow Fan 845
15–4 CFD with Heat Transfer 853 Temperature Rise through a Cross-Flow Heat Exchanger 853 Cooling of an Array of Integrated Circuit Chips
855
15–5 Compressible Flow CFD Calculations 860 Compressible Flow through a Converging–Diverging Nozzle 861 Oblique Shocks over a Wedge 865
15–6 Open-Channel Flow CFD Calculations 866 Flow over a Bump on the Bottom of a Channel 867 Flow through a Sluice Gate (Hydraulic Jump) 868
Application Spotlight: A Virtual Stomach 869 Summary 870 References and Suggested Reading Problems 871
870
Properties of the Atmosphere at High Altitude 897 FIGURE A–12 The Moody Chart for the Friction Factor for Fully Developed Flow in Circular Pipes 898 TABLE A–13 One-dimensional isentropic compressible flow functions for an ideal gas with k 1.4 899 TABLE A–14 One-dimensional normal shock functions for an ideal gas with k 1.4 900 TABLE A–15 Rayleigh flow functions for an ideal gas with k 1.4 901 TABLE A–16 Fanno flow functions for an ideal gas with k 1.4 902
APPENDIX
PROPERTY TABLES AND CHARTS (ENGLISH UNITS) 903 TABLE A–1E
APPENDIX
1
PROPERTY TABLES AND CHARTS (SI UNITS) 885 TABLE A–1
TABLE A–2 TABLE A–3 TABLE A–4 TABLE A–5 TABLE A–6 TABLE A–7 TABLE A–8 TABLE A–9 TABLE A–10
Molar Mass, Gas Constant, and Ideal-Gas Specfic Heats of Some Substances 886 Boiling and Freezing Point Properties 887 Properties of Saturated Water 888 Properties of Saturated Refrigerant-134a 889 Properties of Saturated Ammonia 890 Properties of Saturated Propane 891 Properties of Liquids 892 Properties of Liquid Metals 893 Properties of Air at 1 atm Pressure 894 Properties of Gases at 1 atm Pressure 895
2
Molar Mass, Gas Constant, and Ideal-Gas Specific Heats of Some Substances 904 TABLE A–2E Boiling and Freezing Point Properties 905 TABLE A–3E Properties of Saturated Water 906 TABLE A–4E Properties of Saturated Refrigerant-134a 907 TABLE A–5E Properties of Saturated Ammonia 908 TABLE A–6E Properties of Saturated Propane 909 TABLE A–7E Properties of Liquids 910 TABLE A–8E Properties of Liquid Metals 911 TABLE A–9E Properties of Air at 1 atm Pressure 912 TABLE A–10E Properties of Gases at 1 atm Pressure 913 TABLE A–11E Properties of the Atmosphere at High Altitude 915 Glossary 917 Index 931
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P R E FAC E BACKGROUND Fluid mechanics is an exciting and fascinating subject with unlimited practical applications ranging from microscopic biological systems to automobiles, airplanes, and spacecraft propulsion. Yet fluid mechanics has historically been one of the most challenging subjects for undergraduate students. Unlike earlier freshman- and sophomore-level subjects such as physics, chemistry, and engineering mechanics, where students often learn equations and then “plug and chug” on their calculators, proper analysis of a problem in fluid mechanics requires much more. Oftentimes, students must first assess the problem, make and justify assumptions and/or approximations, apply the relevant physical laws in their proper forms, and solve the resulting equations before ever plugging any numbers into their calculators. Many problems in fluid mechanics require more than just knowledge of the subject, but also physical intuition and experience. Our hope is that this book, through its careful explanations of concepts and its use of numerous practical examples, sketches, figures, and photographs, bridges the gap between knowledge and proper application of that knowledge. Fluid mechanics is a mature subject; the basic equations and approximations are well established and can be found in numerous introductory fluid mechanics books. The books are distinguished from one another in the way the material is presented. An accessible fluid mechanics book should present the material in a progressive order from simple to more difficult, building each chapter upon foundations laid down in previous chapters. In this way, even the traditionally challenging aspects of fluid mechanics can be learned effectively. Fluid mechanics is by its very nature a highly visual subject, and students learn more readily by visual stimulation. It is therefore imperative that a good fluid mechanics book also provide quality figures, photographs, and visual aids that help to explain the significance and meaning of the mathematical expressions.
OBJECTIVES This book is intended for use as a textbook in the first fluid mechanics course for undergraduate engineering students in their junior or senior year. Students are assumed to have an adequate background in calculus, physics, engineering mechanics, and thermodynamics. The objectives of this text are • To cover the basic principles and equations of fluid mechanics • To present numerous and diverse real-world engineering examples to give students a feel for how fluid mechanics is applied in engineering practice • To develop an intuitive understanding of fluid mechanics by emphasizing the physics, and by supplying attractive figures and visual aids to reinforce the physics
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xvi FLUID MECHANICS
The text contains sufficient material to give instructors flexibility as to which topics to emphasize. For example, aeronautics and aerospace engineering instructors may emphasize potential flow, drag and lift, compressible flow, turbomachinery, and CFD, while mechanical and civil engineering instructors may choose to emphasize pipe flows and open-channel flows, respectively. The book has been written with enough breadth of coverage that it can be used for a two-course sequence in fluid mechanics if desired.
PHILOSOPHY AND GOAL We have adopted the same philosophy as that of the texts Thermodynamics: An Engineering Approach by Y. A. Çengel and M. A. Boles, Heat Transfer: A Practical Approach by Y. A. Çengel, and Fundamentals of Thermal-Fluid Sciences by Y. A. Çengel and R. H. Turner, all published by McGraw-Hill. Namely, our goal is to offer an engineering textbook that • Communicates directly to the minds of tomorrow’s engineers in a simple yet precise manner • Leads students toward a clear understanding and firm grasp of the basic principles of fluid mechanics • Encourages creative thinking and development of a deeper understanding and intuitive feel for fluid mechanics • Is read by students with interest and enthusiasm rather than merely as an aid to solve problems It is our philosophy that the best way to learn is by practice. Therefore, special effort is made throughout the book to reinforce material that was presented earlier (both earlier in the chapter and in previous chapters). For example, many of the illustrated example problems and end-of-chapter problems are comprehensive, forcing the student to review concepts learned in previous chapters. Throughout the book, we show examples generated by computational fluid dynamics (CFD), and we provide an introductory chapter on CFD. Our goal is not to teach details about numerical algorithms associated with CFD—this is more properly presented in a separate course, typically at the graduate level. Rather, it is our intent to introduce undergraduate students to the capabilities and limitations of CFD as an engineering tool. We use CFD solutions in much the same way as we use experimental results from a wind tunnel test, i.e., to reinforce understanding of the physics of fluid flows and to provide quality flow visualizations that help to explain fluid behavior.
C O N T E N T A N D O R G A N I Z AT I O N This book is organized into 15 chapters beginning with fundamental concepts of fluids and fluid flows and ending with an introduction to computational fluid dynamics, the application of which is rapidly becoming more commonplace, even at the undergraduate level. • Chapter 1 provides a basic introduction to fluids, classifications of fluid flow, control volume versus system formulations, dimensions, units, significant digits, and problem-solving techniques.
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xvii PREFACE
• Chapter 2 is devoted to fluid properties such as density, vapor pressure, specific heats, viscosity, and surface tension. • Chapter 3 deals with fluid statics and pressure, including manometers and barometers, hydrostatic forces on submerged surfaces, buoyancy and stability, and fluids in rigid-body motion. • Chapter 4 covers topics related to fluid kinematics, such as the differences between Lagrangian and Eulerian descriptions of fluid flows, flow patterns, flow visualization, vorticity and rotationality, and the Reynolds transport theorem. • Chapter 5 introduces the fundamental conservation laws of mass, momentum, and energy, with emphasis on the proper use of the mass, Bernoulli, and energy equations and the engineering applications of these equations. • Chapter 6 applies the Reynolds transport theorem to linear momentum and angular momentum and emphasizes practical engineering applications of the finite control volume momentum analysis. • Chapter 7 reinforces the concept of dimensional homogeneity and introduces the Buckingham Pi theorem of dimensional analysis, dynamic similarity, and the method of repeating variables—material that is useful throughout the rest of the book and in many disciplines in science and engineering. • Chapter 8 is devoted to flow in pipes and ducts. We discuss the differences between laminar and turbulent flow, friction losses in pipes and ducts, and minor losses in piping networks. We also explain how to properly select a pump or fan to match a piping network. Finally, we discuss various experimental devices that are used to measure flow rate and velocity. • Chapter 9 deals with differential analysis of fluid flow and includes derivation and application of the continuity equation, the Cauchy equation, and the Navier–Stokes equation. We also introduce the stream function and describe its usefulness in analysis of fluid flows. • Chapter 10 discusses several approximations of the Navier–Stokes equations and provides example solutions for each approximation, including creeping flow, inviscid flow, irrotational (potential) flow, and boundary layers. • Chapter 11 covers forces on bodies (drag and lift), explaining the distinction between friction and pressure drag, and providing drag coefficients for many common geometries. This chapter emphasizes the practical application of wind tunnel measurements coupled with dynamic similarity and dimensional analysis concepts introduced earlier in Chapter 7. • Chapter 12 extends fluid flow analysis to compressible flow, where the behavior of gases is greatly affected by the Mach number, and the concepts of expansion waves, normal and oblique shock waves, and choked flow are introduced. • Chapter 13 deals with open-channel flow and some of the unique features associated with the flow of liquids with a free surface, such as surface waves and hydraulic jumps.
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• Chapter 14 examines turbomachinery in more detail, including pumps, fans, and turbines. An emphasis is placed on how pumps and turbines work, rather than on their detailed design. We also discuss overall pump and turbine design, based on dynamic similarity laws and simplified velocity vector analyses. • Chapter 15 describes the fundamental concepts of computational fluid dynamics (CFD) and shows students how to use commercial CFD codes as a tool to solve complex fluid mechanics problems. We emphasize the application of CFD rather than the algorithms used in CFD codes. Each chapter contains a large number of end-of-chapter homework problems suitable for use by instructors. Most of the problems that involve calculations are in SI units, but approximately 20 percent are written in English units. Finally, a comprehensive set of appendices is provided, giving the thermodynamic and fluid properties of several materials, not just air and water as in most introductory fluids texts. Many of the end-of-chapter problems require use of the properties found in these appendices.
LEARNING TOOLS EMPHASIS ON PHYSICS A distinctive feature of this book is its emphasis on the physical aspects of the subject matter in addition to mathematical representations and manipulations. The authors believe that the emphasis in undergraduate education should remain on developing a sense of underlying physical mechanisms and a mastery of solving practical problems that an engineer is likely to face in the real world. Developing an intuitive understanding should also make the course a more motivating and worthwhile experience for the students.
EFFECTIVE USE OF ASSOCIATION An observant mind should have no difficulty understanding engineering sciences. After all, the principles of engineering sciences are based on our everyday experiences and experimental observations. Therefore, a physical, intuitive approach is used throughout this text. Frequently, parallels are drawn between the subject matter and students’ everyday experiences so that they can relate the subject matter to what they already know.
SELF-INSTRUCTING The material in the text is introduced at a level that an average student can follow comfortably. It speaks to students, not over students. In fact, it is selfinstructive. Noting that the principles of science are based on experimental observations, most of the derivations in this text are largely based on physical arguments, and thus they are easy to follow and understand.
EXTENSIVE USE OF ARTWORK Figures are important learning tools that help the students “get the picture,” and the text makes effective use of graphics. It contains more figures and illustrations than any other book in this category. Figures attract attention and stimulate curiosity and interest. Most of the figures in this text are intended to serve as a means of emphasizing some key concepts that would otherwise go unnoticed; some serve as page summaries.
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CHAPTER OPENERS AND SUMMARIES Each chapter begins with an overview of the material to be covered. A summary is included at the end of each chapter, providing a quick review of basic concepts and important relations, and pointing out the relevance of the material.
NUMEROUS WORKED-OUT EXAMPLES WITH A SYSTEMATIC SOLUTIONS PROCEDURE Each chapter contains several worked-out examples that clarify the material and illustrate the use of the basic principles. An intuitive and systematic approach is used in the solution of the example problems, while maintaining an informal conversational style. The problem is first stated, and the objectives are identified. The assumptions are then stated, together with their justifications. The properties needed to solve the problem are listed separately. Numerical values are used together with their units to emphasize that numbers without units are meaningless, and unit manipulations are as important as manipulating the numerical values with a calculator. The significance of the findings is discussed following the solutions. This approach is also used consistently in the solutions presented in the instructor’s solutions manual.
A WEALTH OF REALISTIC END-OF-CHAPTER PROBLEMS The end-of-chapter problems are grouped under specific topics to make problem selection easier for both instructors and students. Within each group of problems are Concept Questions, indicated by “C,” to check the students’ level of understanding of basic concepts. The problems under Review Problems are more comprehensive in nature and are not directly tied to any specific section of a chapter – in some cases they require review of material learned in previous chapters. Problems designated as Design and Essay are intended to encourage students to make engineering judgments, to conduct independent exploration of topics of interest, and to communicate their findings in a professional manner. Problems designated by an “E” are in English units, and SI users can ignore them. Problems with the are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text. Several economics- and safety-related problems are incorporated throughout to enhance cost and safety awareness among engineering students. Answers to selected problems are listed immediately following the problem for convenience to students.
USE OF COMMON NOTATION The use of different notation for the same quantities in different engineering courses has long been a source of discontent and confusion. A student taking both fluid mechanics and heat transfer, for example, has to use the notation Q for volume flow rate in one course, and for heat transfer in the other. The need to unify notation in engineering education has often been raised, even in some reports of conferences sponsored by the National Science Foundation through Foundation Coalitions, but little effort has been made to date in this regard. For example, refer to the final report of the “Mini-Conference on Energy Stem Innovations, May 28 and 29, 2003, University of Wisconsin.” In this text we made a conscious effort to minimize this conflict by adopting the familiar
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. thermodynamic notation V for volume flow rate, thus reserving the notation Q for heat transfer. Also, we consistently use an overdot to denote time rate. We think that both students and instructors will appreciate this effort to promote a common notation.
A CHOICE OF SI ALONE OR SI/ENGLISH UNITS In recognition of the fact that English units are still widely used in some industries, both SI and English units are used in this text, with an emphasis on SI. The material in this text can be covered using combined SI/English units or SI units alone, depending on the preference of the instructor. The property tables and charts in the appendices are presented in both units, except the ones that involve dimensionless quantities. Problems, tables, and charts in English units are designated by “E” after the number for easy recognition, and they can be ignored easily by the SI users.
COMBINED COVERAGE OF BERNOULLI AND ENERGY EQUATIONS The Bernoulli equation is one of the most frequently used equations in fluid mechanics, but it is also one of the most misused. Therefore, it is important to emphasize the limitations on the use of this idealized equation and to show how to properly account for imperfections and irreversible losses. In Chapter 5, we do this by introducing the energy equation right after the Bernoulli equation and demonstrating how the solutions of many practical engineering problems differ from those obtained using the Bernoulli equation. This helps students develop a realistic view of the Bernoulli equation.
A SEPARATE CHAPTER ON CFD Commercial Computational Fluid Dynamics (CFD) codes are widely used in engineering practice in the design and analysis of flow systems, and it has become exceedingly important for engineers to have a solid understanding of the fundamental aspects, capabilities, and limitations of CFD. Recognizing that most undergraduate engineering curriculums do not have room for a full course on CFD, a separate chapter is included here to make up for this deficiency and to equip students with an adequate background on the strengths and weaknesses of CFD.
APPLICATION SPOTLIGHTS Throughout the book are highlighted examples called Application Spotlights where a real-world application of fluid mechanics is shown. A unique feature of these special examples is that they are written by guest authors. The Application Spotlights are designed to show students how fluid mechanics has diverse applications in a wide variety of fields. They also include eye-catching photographs from the guest authors’ research.
GLOSSARY OF FLUID MECHANICS TERMS Throughout the chapters, when an important key term or concept is introduced and defined, it appears in black boldface type. Fundamental fluid mechanics terms and concepts appear in blue boldface type, and these fundamental terms also appear in a comprehensive end-of-book glossary developed by Professor James Brasseur of The Pennsylvania State University. This unique glossary is an excellent learning and review tool for students as they move forward in
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their study of fluid mechanics. In addition, students can test their knowledge of these fundamental terms by using the interactive flash cards and other resources located on our accompanying website (www.mhhe.com/cengel).
CONVERSION FACTORS Frequently used conversion factors, physical constants, and frequently used properties of air and water at 20°C and atmospheric pressure are listed on the front inner cover pages of the text for easy reference.
NOMENCLATURE A list of the major symbols, subscripts, and superscripts used in the text are listed on the inside back cover pages of the text for easy reference.
SUPPLEMENTS These supplements are available to adopters of the book:
STUDENT RESOURCES DVD Packaged free with every new copy of the text, this DVD provides a wealth of resources for students including Fluid Mechanics Videos, a CFD Animations Library, and EES Software.
ONLINE LEARNING CENTER Web support is provided for the book on our Online Learning Center at www.mhhe.com/cengel. Visit this robust site for book and supplement information, errata, author information, and further resources for instructors and students.
ENGINEERING EQUATION SOLVER (EES) Developed by Sanford Klein and William Beckman from the University of Wisconsin–Madison, this software combines equation-solving capability and engineering property data. EES can do optimization, parametric analysis, and linear and nonlinear regression, and provides publication-quality plotting capabilities. Thermodynamics and transport properties for air, water, and many other fluids are built-in and EES allows the user to enter property data or functional relationships.
FLUENT FLOWLAB® SOFTWARE AND TEMPLATES As an integral part of Chapter 15, “Introduction to Computational Fluid Dynamics,” we provide access to a student-friendly CFD software package developed by Fluent Inc. In addition, we provide over 40 FLUENT FLOWLAB templates to complement the end-of-chapter problems in Chapter 15. These problems and templates are unique in that they are designed with both a fluid mechanics learning objective and a CFD learning objective in mind.
INSTRUCTOR’S RESOURCE CD-ROM (AVAILABLE TO INSTRUCTORS ONLY) This CD, available to instructors only, offers a wide range of classroom preparation and presentation resources including an electronic solutions manual with PDF files by chapter, all text chapters and appendices as downloadable PDF files, and all text figures in JPEG format.
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COSMOS CD-ROM (AVAILABLE TO INSTRUCTORS ONLY) This CD, available to instructors only, provides electronic solutions delivered via our database management tool. McGraw-Hill’s COSMOS allows instructors to streamline the creation of assignments, quizzes, and tests by using problems and solutions from the textbook—as well as their own custom material.
ACKNOWLEDGMENTS The authors would like to acknowledge with appreciation the numerous and valuable comments, suggestions, constructive criticisms, and praise from the following evaluators and reviewers: Mohammad Ali Kettering University
Darryl Alofs University of Missouri, Rolla
Farrukh Alvi Florida A & M University & Florida State University
Ryoichi Amano
Soyoung Cha University of Illinois at Chicago
Tiao Chang Ohio University
Young Cho Drexel University
Po-Ya (Abel) Chuang The Pennsylvania State University
University of Wisconsin–Milwaukee
Michael Amitay
William H. Colwill American Hydro Corporation
Rensselaer Polytechnic Institute
T. P. Ashokbabu
A. Terrence Conlisk Jr. The Ohio State University
National Institute of Technology, India
Idirb Azouz
Daniel Cox Texas A&M University
Southern Utah University
Kenneth S. Ball
John Crepeau University of Idaho
University of Texas at Austin
James G. Brasseur
Jie Cui Tennessee Technological University
The Pennsylvania State University
Glenn Brown
Lisa Davids Embry-Riddle Aeronautical University
Oklahoma State University
John Callister
Jerry Drummond The University of Akron
Cornell University
Frederick Carranti
Dwayne Edwards University of Kentucky
Syracuse University
Kevin W. Cassel
Richard Figliola Clemson University
Illinois Institute of Technology
Haris Catrakis
Charles Forsberg Hofstra University
University of California, Irvine
Louis N. Cattafesta III University of Florida
Fred K. Forster University of Washington
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Rong Gan The University of Oklahoma
Philip Gerhart University of Evansville
Fred Gessner University of Washington
Sam Han Tennessee Technological University
Mark J. Holowach Ballston Spa, NY
Neal Houze Purdue University
Barbara Hutchings Fluent Incorporated
Niu Jianlei Hong Kong Polytechnic University, Hong Kong
David Johnson
James A. Liburdy Oregon State University
Chao-An Lin National Tsing Hua University, Taiwan
Kraemer Luks The University of Tulsa
G. Mahinthakumar North Carolina State University
Saeed Manafzadeh University of Illinois at Chicago
Daniel Maynes Brigham Young University
James M. McDonough University of Kentucky
Richard S. Miller Clemson University
Shane Moeykens Fluent Incorporated
University of Waterloo
Matthew Jones
Joseph Morrison NASA Langley Research Center
Brigham Young University
Zbigniew J. Kabala
Karim Nasr Kettering University
Duke University
Fazal Kauser California State Polytechnic University, Pomona
Pirouz Kavehpour University of California, Los Angeles
Jacob Kazakia
C. O. Ng University of Hong Kong, Hong Kong
Wing Ng Virginia Polytechnic Institute
Tay Seow Ngie Nanyang Technological University, Singapore
Lehigh University
Richard Keane University of Illinois at Urbana–Champaign
Jamil Khan
John Nicklow Southern Illinois University at Carbondale
Nagy Nosseir San Diego State University
University of South Carolina
N. Nirmala Khandan
Emmanuel Nzewi North Carolina A&T State University
New Mexico State University
Jeyhoon Khodadadi
Ali Ogut Rochester Institute of Technology
Auburn University
Subha Kumpaty Milwaukee School of Engineering
Michael Olsen Iowa State University
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Roger Pawlowski Lawrence Technological University
Bryan Pearce The University of Maine
Blair Perot University of Massachusetts Amherst
Alexander Povitsky The University of Akron
Guy Riefler Ohio University
Kurt Rosentrater Northern Illinois University
Mark Stone Washington State University
Chelakara Subramanian Florida Institute of Technology
Constantine Tarawneh The University of Texas–Pan American
Sahnaz Tigrek Middle East Technical University
Hsu Chin Tsau Hong Kong University of Science and Technology, Hong Kong M.
Erol Ulucakli Lafayette College
Subrata Roy Kettering University
Oleg Vasilyev University of Missouri
Joseph Sai Texas A&M University–Kingsville
Zhi Jian Wang Michigan State University
Gregory Selby Old Dominion University
Gary S. Settles The Pennsylvania State University
Winoto SH National University of Singapore, Singapore
Timothy Wei Rutgers, The State University of New Jersey
Minami Yoda Georgia Institute of Technology
Mohd Zamri Yusoff Universiti Tenaga Nasional, Malaysia
Muhammad Sharif The University of Alabama
The authors also acknowledge the guest authors who contributed photographs and write-ups for the Application Spotlights: Michael L. Billet The Pennsylvania State University
James G. Brasseur The Pennsylvania State University
Werner J. A. Dahm University of Michigan
Brian Daniels Oregon State University
Michael Dickinson California Institute of Technology
Gerald C. Lauchle The Pennsylvania State University
James A. Liburdy Oregon State University
Anupam Pal The Pennsylvania State University
Ganesh Raman Illinois Institute of Technology
Gary S. Settles The Pennsylvania State University
Lorenz Sigurdson University of Alberta
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Special thanks go to Professor Gary Settles and his associates at Penn State (Lori Dodson-Dreibelbis, J. D. Miller, and Gabrielle Tremblay) for creating the exciting narrated video clips that are found on the DVD that accompanies this book. Similarly, the authors acknowledge several people at Fluent Inc., who helped to make available the wonderful CFD animations that are also found on the DVD and the FLUENT FLOWLAB templates that are available for downloading from the book’s website: Shane Moeykens, Barbara Hutchings, Liz Marshall, Ashish Kulkarni, Ajay Parihar, and R. Murali Krishnan. The authors also thank Professor James Brasseur of Penn State for creating the precise glossary of fluid mechanics terms, Professor Glenn Brown of Oklahoma State for providing many items of historical interest throughout the text, Professor Mehmet Kanoglu of Gaziantep University for preparing the solutions of EES problems, and Professor Tahsin Engin of Sakarya University for contributing several end-of-chapter problems. Finally, special thanks must go to our families, especially our wives, Zehra Çengel and Suzanne Cimbala, for their continued patience, understanding, and support throughout the preparation of this book, which involved many long hours when they had to handle family concerns on their own because their husbands’ faces were glued to a computer screen. Yunus A. Çengel John M. Cimbala
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CHAPTER
1
INTRODUCTION AND BASIC CONCEPTS
I
n this introductory chapter, we present the basic concepts commonly used in the analysis of fluid flow. We start this chapter with a discussion of the phases of matter and the numerous ways of classification of fluid flow, such as viscous versus inviscid regions of flow, internal versus external flow, compressible versus incompressible flow, laminar versus turbulent flow, natural versus forced flow, and steady versus unsteady flow. We also discuss the no-slip condition at solid–fluid interfaces and present a brief history of the development of fluid mechanics. After presenting the concepts of system and control volume, we review the unit systems that will be used. We then discuss how mathematical models for engineering problems are prepared and how to interpret the results obtained from the analysis of such models. This is followed by a presentation of an intuitive systematic problem-solving technique that can be used as a model in solving engineering problems. Finally, we discuss accuracy, precision, and significant digits in engineering measurements and calculations.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
Understand the basic concepts of fluid mechanics and recognize the various types of fluid flow problems encountered in practice Model engineering problems and solve them in a systematic manner Have a working knowledge of accuracy, precision, and significant digits, and recognize the importance of dimensional homogeneity in engineering calculations
1
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1–1
FIGURE 1–1 Fluid mechanics deals with liquids and gases in motion or at rest. © Vol. 16/Photo Disc.
■
INTRODUCTION
Mechanics is the oldest physical science that deals with both stationary and moving bodies under the influence of forces. The branch of mechanics that deals with bodies at rest is called statics, while the branch that deals with bodies in motion is called dynamics. The subcategory fluid mechanics is defined as the science that deals with the behavior of fluids at rest (fluid statics) or in motion (fluid dynamics), and the interaction of fluids with solids or other fluids at the boundaries. Fluid mechanics is also referred to as fluid dynamics by considering fluids at rest as a special case of motion with zero velocity (Fig. 1–1). Fluid mechanics itself is also divided into several categories. The study of the motion of fluids that are practically incompressible (such as liquids, especially water, and gases at low speeds) is usually referred to as hydrodynamics. A subcategory of hydrodynamics is hydraulics, which deals with liquid flows in pipes and open channels. Gas dynamics deals with the flow of fluids that undergo significant density changes, such as the flow of gases through nozzles at high speeds. The category aerodynamics deals with the flow of gases (especially air) over bodies such as aircraft, rockets, and automobiles at high or low speeds. Some other specialized categories such as meteorology, oceanography, and hydrology deal with naturally occurring flows.
What Is a Fluid?
Contact area, A a
Shear stress t = F/A
Force, F
Deformed rubber
Shear strain, a
FIGURE 1–2 Deformation of a rubber eraser placed between two parallel plates under the influence of a shear force.
You will recall from physics that a substance exists in three primary phases: solid, liquid, and gas. (At very high temperatures, it also exists as plasma.) A substance in the liquid or gas phase is referred to as a fluid. Distinction between a solid and a fluid is made on the basis of the substance’s ability to resist an applied shear (or tangential) stress that tends to change its shape. A solid can resist an applied shear stress by deforming, whereas a fluid deforms continuously under the influence of shear stress, no matter how small. In solids stress is proportional to strain, but in fluids stress is proportional to strain rate. When a constant shear force is applied, a solid eventually stops deforming, at some fixed strain angle, whereas a fluid never stops deforming and approaches a certain rate of strain. Consider a rectangular rubber block tightly placed between two plates. As the upper plate is pulled with a force F while the lower plate is held fixed, the rubber block deforms, as shown in Fig. 1–2. The angle of deformation a (called the shear strain or angular displacement) increases in proportion to the applied force F. Assuming there is no slip between the rubber and the plates, the upper surface of the rubber is displaced by an amount equal to the displacement of the upper plate while the lower surface remains stationary. In equilibrium, the net force acting on the plate in the horizontal direction must be zero, and thus a force equal and opposite to F must be acting on the plate. This opposing force that develops at the plate–rubber interface due to friction is expressed as F ! tA, where t is the shear stress and A is the contact area between the upper plate and the rubber. When the force is removed, the rubber returns to its original position. This phenomenon would also be observed with other solids such as a steel block provided that the applied force does not exceed the elastic range. If this experiment were repeated with a fluid (with two large parallel plates placed in a large body of water, for example), the fluid layer in contact with the upper plate would
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move with the plate continuously at the velocity of the plate no matter how small the force F is. The fluid velocity decreases with depth because of friction between fluid layers, reaching zero at the lower plate. You will recall from statics that stress is defined as force per unit area and is determined by dividing the force by the area upon which it acts. The normal component of the force acting on a surface per unit area is called the normal stress, and the tangential component of a force acting on a surface per unit area is called shear stress (Fig. 1–3). In a fluid at rest, the normal stress is called pressure. The supporting walls of a fluid eliminate shear stress, and thus a fluid at rest is at a state of zero shear stress. When the walls are removed or a liquid container is tilted, a shear develops and the liquid splashes or moves to attain a horizontal free surface. In a liquid, chunks of molecules can move relative to each other, but the volume remains relatively constant because of the strong cohesive forces between the molecules. As a result, a liquid takes the shape of the container it is in, and it forms a free surface in a larger container in a gravitational field. A gas, on the other hand, expands until it encounters the walls of the container and fills the entire available space. This is because the gas molecules are widely spaced, and the cohesive forces between them are very small. Unlike liquids, gases cannot form a free surface (Fig. 1–4). Although solids and fluids are easily distinguished in most cases, this distinction is not so clear in some borderline cases. For example, asphalt appears and behaves as a solid since it resists shear stress for short periods of time. But it deforms slowly and behaves like a fluid when these forces are exerted for extended periods of time. Some plastics, lead, and slurry mixtures exhibit similar behavior. Such borderline cases are beyond the scope of this text. The fluids we will deal with in this text will be clearly recognizable as fluids. Intermolecular bonds are strongest in solids and weakest in gases. One reason is that molecules in solids are closely packed together, whereas in gases they are separated by relatively large distances (Fig. 1–5). The molecules in a solid are arranged in a pattern that is repeated throughout. Because of the small distances between molecules in a solid, the attractive forces of molecules on each other are large and keep the molecules at
(a)
(b)
Normal to surface Force acting F on area dA
Fn
dA
Tangent to surface
Ft
Normal stress: s ! Shear stress: t !
Fn dA Ft dA
FIGURE 1–3 The normal stress and shear stress at the surface of a fluid element. For fluids at rest, the shear stress is zero and pressure is the only normal stress.
Free surface
Liquid
Gas
FIGURE 1–4 Unlike a liquid, a gas does not form a free surface, and it expands to fill the entire available space.
(c)
FIGURE 1–5 The arrangement of atoms in different phases: (a) molecules are at relatively fixed positions in a solid, (b) groups of molecules move about each other in the liquid phase, and (c) molecules move about at random in the gas phase.
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Pressure gage
FIGURE 1–6 On a microscopic scale, pressure is determined by the interaction of individual gas molecules. However, we can measure the pressure on a macroscopic scale with a pressure gage.
fixed positions. The molecular spacing in the liquid phase is not much different from that of the solid phase, except the molecules are no longer at fixed positions relative to each other and they can rotate and translate freely. In a liquid, the intermolecular forces are weaker relative to solids, but still strong compared with gases. The distances between molecules generally increase slightly as a solid turns liquid, with water being a notable exception. In the gas phase, the molecules are far apart from each other, and a molecular order is nonexistent. Gas molecules move about at random, continually colliding with each other and the walls of the container in which they are contained. Particularly at low densities, the intermolecular forces are very small, and collisions are the only mode of interaction between the molecules. Molecules in the gas phase are at a considerably higher energy level than they are in the liquid or solid phase. Therefore, the gas must release a large amount of its energy before it can condense or freeze. Gas and vapor are often used as synonymous words. The vapor phase of a substance is customarily called a gas when it is above the critical temperature. Vapor usually implies a gas that is not far from a state of condensation. Any practical fluid system consists of a large number of molecules, and the properties of the system naturally depend on the behavior of these molecules. For example, the pressure of a gas in a container is the result of momentum transfer between the molecules and the walls of the container. However, one does not need to know the behavior of the gas molecules to determine the pressure in the container. It would be sufficient to attach a pressure gage to the container (Fig. 1–6). This macroscopic or classical approach does not require a knowledge of the behavior of individual molecules and provides a direct and easy way to the solution of engineering problems. The more elaborate microscopic or statistical approach, based on the average behavior of large groups of individual molecules, is rather involved and is used in this text only in the supporting role.
Application Areas of Fluid Mechanics Fluid mechanics is widely used both in everyday activities and in the design of modern engineering systems from vacuum cleaners to supersonic aircraft. Therefore, it is important to develop a good understanding of the basic principles of fluid mechanics. To begin with, fluid mechanics plays a vital role in the human body. The heart is constantly pumping blood to all parts of the human body through the arteries and veins, and the lungs are the sites of airflow in alternating directions. Needless to say, all artificial hearts, breathing machines, and dialysis systems are designed using fluid dynamics. An ordinary house is, in some respects, an exhibition hall filled with applications of fluid mechanics. The piping systems for cold water, natural gas, and sewage for an individual house and the entire city are designed primarily on the basis of fluid mechanics. The same is also true for the piping and ducting network of heating and air-conditioning systems. A refrigerator involves tubes through which the refrigerant flows, a compressor that pressurizes the refrigerant, and two heat exchangers where the refrigerant absorbs and rejects heat. Fluid mechanics plays a major role in the design of all these components. Even the operation of ordinary faucets is based on fluid mechanics. We can also see numerous applications of fluid mechanics in an automobile. All components associated with the transportation of the fuel from the
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fuel tank to the cylinders—the fuel line, fuel pump, fuel injectors, or carburetors—as well as the mixing of the fuel and the air in the cylinders and the purging of combustion gases in exhaust pipes are analyzed using fluid mechanics. Fluid mechanics is also used in the design of the heating and air-conditioning system, the hydraulic brakes, the power steering, automatic transmission, and lubrication systems, the cooling system of the engine block including the radiator and the water pump, and even the tires. The sleek streamlined shape of recent model cars is the result of efforts to minimize drag by using extensive analysis of flow over surfaces. On a broader scale, fluid mechanics plays a major part in the design and analysis of aircraft, boats, submarines, rockets, jet engines, wind turbines, biomedical devices, the cooling of electronic components, and the transportation of water, crude oil, and natural gas. It is also considered in the design of buildings, bridges, and even billboards to make sure that the structures can withstand wind loading. Numerous natural phenomena such as the rain cycle, weather patterns, the rise of ground water to the top of trees, winds, ocean waves, and currents in large water bodies are also governed by the principles of fluid mechanics (Fig. 1–7).
Natural flows and weather
Boats
Aircraft and spacecraft
© Vol. 16/Photo Disc.
© Vol. 5/Photo Disc.
© Vol. 1/Photo Disc.
Power plants
Human body
Cars
© Vol. 57/Photo Disc.
© Vol. 110/Photo Disc.
Photo by John M. Cimbala.
Wind turbines
Piping and plumbing systems
Industrial applications
© Vol. 17/Photo Disc.
Photo by John M. Cimbala.
Courtesy UMDE Engineering, Contracting, and Trading. Used by permission.
FIGURE 1–7 Some application areas of fluid mechanics.
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1–2
FIGURE 1–8 The development of a velocity profile due to the no-slip condition as a fluid flows over a blunt nose. “Hunter Rouse: Laminar and Turbulent Flow Film.” Copyright IIHR-Hydroscience & Engineering, The University of Iowa. Used by permission.
Uniform approach velocity, V
Relative velocities of fluid layers Zero velocity at the surface
Plate
FIGURE 1–9 A fluid flowing over a stationary surface comes to a complete stop at the surface because of the no-slip condition.
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THE NO-SLIP CONDITION
Fluid flow is often confined by solid surfaces, and it is important to understand how the presence of solid surfaces affects fluid flow. We know that water in a river cannot flow through large rocks, and goes around them. That is, the water velocity normal to the rock surface must be zero, and water approaching the surface normally comes to a complete stop at the surface. What is not so obvious is that water approaching the rock at any angle also comes to a complete stop at the rock surface, and thus the tangential velocity of water at the surface is also zero. Consider the flow of a fluid in a stationary pipe or over a solid surface that is nonporous (i.e., impermeable to the fluid). All experimental observations indicate that a fluid in motion comes to a complete stop at the surface and assumes a zero velocity relative to the surface. That is, a fluid in direct contact with a solid “sticks” to the surface due to viscous effects, and there is no slip. This is known as the no-slip condition. The photo in Fig. 1–8 obtained from a video clip clearly shows the evolution of a velocity gradient as a result of the fluid sticking to the surface of a blunt nose. The layer that sticks to the surface slows the adjacent fluid layer because of viscous forces between the fluid layers, which slows the next layer, and so on. Therefore, the no-slip condition is responsible for the development of the velocity profile. The flow region adjacent to the wall in which the viscous effects (and thus the velocity gradients) are significant is called the boundary layer. The fluid property responsible for the no-slip condition and the development of the boundary layer is viscosity and is discussed in Chap. 2. A fluid layer adjacent to a moving surface has the same velocity as the surface. A consequence of the no-slip condition is that all velocity profiles must have zero values with respect to the surface at the points of contact between a fluid and a solid surface (Fig. 1–9). Another consequence of the no-slip condition is the surface drag, which is the force a fluid exerts on a surface in the flow direction. When a fluid is forced to flow over a curved surface, such as the back side of a cylinder at sufficiently high velocity, the boundary layer can no longer remain attached to the surface, and at some point it separates from the surface—a process called flow separation (Fig. 1–10). We emphasize that the no-slip condition applies everywhere along the surface, even downstream of the separation point. Flow separation is discussed in greater detail in Chap. 10.
Separation point
FIGURE 1–10 Flow separation during flow over a curved surface. From G. M. Homsy et al, “Multi-Media Fluid Mechanics,” Cambridge Univ. Press (2001). ISBN 0-521-78748-3. Reprinted by permission.
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A similar phenomenon occurs for temperature. When two bodies at different temperatures are brought into contact, heat transfer occurs until both bodies assume the same temperature at the points of contact. Therefore, a fluid and a solid surface have the same temperature at the points of contact. This is known as no-temperature-jump condition.
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A BRIEF HISTORY OF FLUID MECHANICS1
One of the first engineering problems humankind faced as cities were developed was the supply of water for domestic use and irrigation of crops. Our urban lifestyles can be retained only with abundant water, and it is clear from archeology that every successful civilization of prehistory invested in the construction and maintenance of water systems. The Roman aqueducts, some of which are still in use, are the best known examples. However, perhaps the most impressive engineering from a technical viewpoint was done at the Hellenistic city of Pergamon in present-day Turkey. There, from 283 to 133 BC, they built a series of pressurized lead and clay pipelines (Fig. 1–11), up to 45 km long that operated at pressures exceeding 1.7 MPa (180 m of head). Unfortunately, the names of almost all these early builders are lost to history. The earliest recognized contribution to fluid mechanics theory was made by the Greek mathematician Archimedes (285–212 BC). He formulated and applied the buoyancy principle in history’s first nondestructive test to determine the gold content of the crown of King Hiero I. The Romans built great aqueducts and educated many conquered people on the benefits of clean water, but overall had a poor understanding of fluids theory. (Perhaps they shouldn’t have killed Archimedes when they sacked Syracuse.) During the Middle Ages the application of fluid machinery slowly but steadily expanded. Elegant piston pumps were developed for dewatering mines, and the watermill and windmill were perfected to grind grain, forge metal, and for other tasks. For the first time in recorded human history significant work was being done without the power of a muscle supplied by a person or animal, and these inventions are generally credited with enabling the later industrial revolution. Again the creators of most of the progress are unknown, but the devices themselves were well documented by several technical writers such as Georgius Agricola (Fig. 1–12). The Renaissance brought continued development of fluid systems and machines, but more importantly, the scientific method was perfected and adopted throughout Europe. Simon Stevin (1548–1617), Galileo Galilei (1564–1642), Edme Mariotte (1620–1684), and Evangelista Torricelli (1608–1647) were among the first to apply the method to fluids as they investigated hydrostatic pressure distributions and vacuums. That work was integrated and refined by the brilliant mathematician, Blaise Pascal (1623– 1662). The Italian monk, Benedetto Castelli (1577–1644) was the first person to publish a statement of the continuity principle for fluids. Besides formulating his equations of motion for solids, Sir Isaac Newton (1643–1727) applied his laws to fluids and explored fluid inertia and resistance, free jets, and viscosity. That effort was built upon by the Swiss Daniel Bernoulli
1
This section is contributed by Professor Glenn Brown of Oklahoma State University.
FIGURE 1–11 Segment of Pergamon pipeline. Each clay pipe section was 13 to 18 cm in diameter. Courtesy Gunther Garbrecht. Used by permission.
FIGURE 1–12 A mine hoist powered by a reversible water wheel. G. Agricola, De Re Metalica, Basel, 1556.
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FIGURE 1–13 The Wright brothers take flight at Kitty Hawk. National Air and Space Museum/ Smithsonian Institution.
(1700–1782) and his associate Leonard Euler (1707–1783). Together, their work defined the energy and momentum equations. Bernoulli’s 1738 classic treatise Hydrodynamica may be considered the first fluid mechanics text. Finally, Jean d’Alembert (1717–1789) developed the idea of velocity and acceleration components, a differential expression of continuity, and his “paradox” of zero resistance to steady uniform motion. The development of fluid mechanics theory up through the end of the eighteenth century had little impact on engineering since fluid properties and parameters were poorly quantified, and most theories were abstractions that could not be quantified for design purposes. That was to change with the development of the French school of engineering led by Riche de Prony (1755–1839). Prony (still known for his brake to measure power) and his associates in Paris at the Ecole Polytechnic and the Ecole Ponts et Chaussees were the first to integrate calculus and scientific theory into the engineering curriculum, which became the model for the rest of the world. (So now you know whom to blame for your painful freshman year.) Antonie Chezy (1718–1798), Louis Navier (1785–1836), Gaspard Coriolis (1792–1843), Henry Darcy (1803–1858), and many other contributors to fluid engineering and theory were students and/or instructors at the schools. By the mid nineteenth century fundamental advances were coming on several fronts. The physician Jean Poiseuille (1799–1869) had accurately measured flow in capillary tubes for multiple fluids, while in Germany Gotthilf Hagen (1797–1884) had differentiated between laminar and turbulent flow in pipes. In England, Lord Osborn Reynolds (1842–1912) continued that work and developed the dimensionless number that bears his name. Similarly, in parallel to the early work of Navier, George Stokes (1819– 1903) completed the general equations of fluid motion with friction that take their names. William Froude (1810–1879) almost single-handedly developed the procedures and proved the value of physical model testing. American expertise had become equal to the Europeans as demonstrated by James Francis’s (1815–1892) and Lester Pelton’s (1829–1908) pioneering work in turbines and Clemens Herschel’s (1842–1930) invention of the Venturi meter. The late nineteenth century was notable for the expansion of fluid theory by Irish and English scientists and engineers, including in addition to Reynolds and Stokes, William Thomson, Lord Kelvin (1824–1907), William Strutt, Lord Rayleigh (1842–1919), and Sir Horace Lamb (1849–1934). These individuals investigated a large number of problems including dimensional analysis, irrotational flow, vortex motion, cavitation, and waves. In a broader sense their work also explored the links between fluid mechanics, thermodynamics, and heat transfer. The dawn of the twentieth century brought two monumental developments. First in 1903, the self-taught Wright brothers (Wilbur, 1867–1912; Orville, 1871–1948) through application of theory and determined experimentation perfected the airplane. Their primitive invention was complete and contained all the major aspects of modern craft (Fig. 1–13). The Navier–Stokes equations were of little use up to this time because they were too difficult to solve. In a pioneering paper in 1904, the German Ludwig Prandtl (1875–1953) showed that fluid flows can be divided into a layer near the walls, the boundary layer, where the friction effects are significant and an outer layer where such effects are negligible and the simplified Euler
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and Bernoulli equations are applicable. His students, Theodore von Kármán (1881–1963), Paul Blasius (1883–1970), Johann Nikuradse (1894–1979), and others, built on that theory in both hydraulic and aerodynamic applications. (During World War II, both sides benefited from the theory as Prandtl remained in Germany while his best student, the Hungarian born Theodore von Kármán, worked in America.) The mid twentieth century could be considered a golden age of fluid mechanics applications. Existing theories were adequate for the tasks at hand, and fluid properties and parameters were well defined. These supported a huge expansion of the aeronautical, chemical, industrial, and water resources sectors; each of which pushed fluid mechanics in new directions. Fluid mechanics research and work in the late twentieth century were dominated by the development of the digital computer in America. The ability to solve large complex problems, such as global climate modeling or to optimize the design of a turbine blade, has provided a benefit to our society that the eighteenth-century developers of fluid mechanics could never have imagined (Fig. 1–14). The principles presented in the following pages have been applied to flows ranging from a moment at the microscopic scale to 50 years of simulation for an entire river basin. It is truly mind-boggling. Where will fluid mechanics go in the twenty-first century? Frankly, even a limited extrapolation beyond the present would be sheer folly. However, if history tells us anything, it is that engineers will be applying what they know to benefit society, researching what they don’t know, and having a great time in the process.
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CLASSIFICATION OF FLUID FLOWS
Earlier we defined fluid mechanics as the science that deals with the behavior of fluids at rest or in motion, and the interaction of fluids with solids or other fluids at the boundaries. There is a wide variety of fluid flow problems encountered in practice, and it is usually convenient to classify them on the basis of some common characteristics to make it feasible to study them in groups. There are many ways to classify fluid flow problems, and here we present some general categories.
Viscous versus Inviscid Regions of Flow
When two fluid layers move relative to each other, a friction force develops between them and the slower layer tries to slow down the faster layer. This internal resistance to flow is quantified by the fluid property viscosity, which is a measure of internal stickiness of the fluid. Viscosity is caused by cohesive forces between the molecules in liquids and by molecular collisions in gases. There is no fluid with zero viscosity, and thus all fluid flows involve viscous effects to some degree. Flows in which the frictional effects are significant are called viscous flows. However, in many flows of practical interest, there are regions (typically regions not close to solid surfaces) where viscous forces are negligibly small compared to inertial or pressure forces. Neglecting the viscous terms in such inviscid flow regions greatly simplifies the analysis without much loss in accuracy. The development of viscous and inviscid regions of flow as a result of inserting a flat plate parallel into a fluid stream of uniform velocity is shown in Fig. 1–15. The fluid sticks to the plate on both sides because of
FIGURE 1–14 The Oklahoma Wind Power Center near Woodward consists of 68 turbines, 1.5 MW each. Courtesy Steve Stadler, Oklahoma Wind Power Initiative. Used by permission.
Inviscid flow region Viscous flow region Inviscid flow region
FIGURE 1–15 The flow of an originally uniform fluid stream over a flat plate, and the regions of viscous flow (next to the plate on both sides) and inviscid flow (away from the plate). Fundamentals of Boundary Layers, National Committee from Fluid Mechanics Films, © Education Development Center.
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the no-slip condition, and the thin boundary layer in which the viscous effects are significant near the plate surface is the viscous flow region. The region of flow on both sides away from the plate and unaffected by the presence of the plate is the inviscid flow region.
Internal versus External Flow
FIGURE 1–16 External flow over a tennis ball, and the turbulent wake region behind. Courtesy NASA and Cislunar Aerospace, Inc.
A fluid flow is classified as being internal or external, depending on whether the fluid is forced to flow in a confined channel or over a surface. The flow of an unbounded fluid over a surface such as a plate, a wire, or a pipe is external flow. The flow in a pipe or duct is internal flow if the fluid is completely bounded by solid surfaces. Water flow in a pipe, for example, is internal flow, and airflow over a ball or over an exposed pipe during a windy day is external flow (Fig. 1–16). The flow of liquids in a duct is called open-channel flow if the duct is only partially filled with the liquid and there is a free surface. The flows of water in rivers and irrigation ditches are examples of such flows. Internal flows are dominated by the influence of viscosity throughout the flow field. In external flows the viscous effects are limited to boundary layers near solid surfaces and to wake regions downstream of bodies.
Compressible versus Incompressible Flow A flow is classified as being compressible or incompressible, depending on the level of variation of density during flow. Incompressibility is an approximation, and a flow is said to be incompressible if the density remains nearly constant throughout. Therefore, the volume of every portion of fluid remains unchanged over the course of its motion when the flow (or the fluid) is incompressible. The densities of liquids are essentially constant, and thus the flow of liquids is typically incompressible. Therefore, liquids are usually referred to as incompressible substances. A pressure of 210 atm, for example, causes the density of liquid water at 1 atm to change by just 1 percent. Gases, on the other hand, are highly compressible. A pressure change of just 0.01 atm, for example, causes a change of 1 percent in the density of atmospheric air. When analyzing rockets, spacecraft, and other systems that involve highspeed gas flows, the flow speed is often expressed in terms of the dimensionless Mach number defined as Ma !
Speed of flow V ! c Speed of sound
where c is the speed of sound whose value is 346 m/s in air at room temperature at sea level. A flow is called sonic when Ma ! 1, subsonic when Ma " 1, supersonic when Ma # 1, and hypersonic when Ma ## 1. Liquid flows are incompressible to a high level of accuracy, but the level of variation in density in gas flows and the consequent level of approximation made when modeling gas flows as incompressible depends on the Mach number. Gas flows can often be approximated as incompressible if the density changes are under about 5 percent, which is usually the case when Ma " 0.3. Therefore, the compressibility effects of air can be neglected at speeds under about 100 m/s. Note that the flow of a gas is not necessarily a compressible flow.
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Small density changes of liquids corresponding to large pressure changes can still have important consequences. The irritating “water hammer” in a water pipe, for example, is caused by the vibrations of the pipe generated by the reflection of pressure waves following the sudden closing of the valves.
Laminar
Laminar versus Turbulent Flow Some flows are smooth and orderly while others are rather chaotic. The highly ordered fluid motion characterized by smooth layers of fluid is called laminar. The word laminar comes from the movement of adjacent fluid particles together in “laminates.” The flow of high-viscosity fluids such as oils at low velocities is typically laminar. The highly disordered fluid motion that typically occurs at high velocities and is characterized by velocity fluctuations is called turbulent (Fig. 1–17). The flow of low-viscosity fluids such as air at high velocities is typically turbulent. The flow regime greatly influences the required power for pumping. A flow that alternates between being laminar and turbulent is called transitional. The experiments conducted by Osborn Reynolds in the 1880s resulted in the establishment of the dimensionless Reynolds number, Re, as the key parameter for the determination of the flow regime in pipes (Chap. 8).
Transitional
Turbulent
FIGURE 1–17 Laminar, transitional, and turbulent flows. Courtesy ONERA, photograph by Werlé.
Natural (or Unforced) versus Forced Flow A fluid flow is said to be natural or forced, depending on how the fluid motion is initiated. In forced flow, a fluid is forced to flow over a surface or in a pipe by external means such as a pump or a fan. In natural flows, any fluid motion is due to natural means such as the buoyancy effect, which manifests itself as the rise of the warmer (and thus lighter) fluid and the fall of cooler (and thus denser) fluid (Fig. 1–18). In solar hot-water systems, for example, the thermosiphoning effect is commonly used to replace pumps by placing the water tank sufficiently above the solar collectors.
Steady versus Unsteady Flow The terms steady and uniform are used frequently in engineering, and thus it is important to have a clear understanding of their meanings. The term steady implies no change at a point with time. The opposite of steady is unsteady. The term uniform implies no change with location over a specified region. These meanings are consistent with their everyday use (steady girlfriend, uniform distribution, etc.). The terms unsteady and transient are often used interchangeably, but these terms are not synonyms. In fluid mechanics, unsteady is the most general term that applies to any flow that is not steady, but transient is typically used for developing flows. When a rocket engine is fired up, for example, there are transient effects (the pressure builds up inside the rocket engine, the flow accelerates, etc.) until the engine settles down and operates steadily. The term periodic refers to the kind of unsteady flow in which the flow oscillates about a steady mean. Many devices such as turbines, compressors, boilers, condensers, and heat exchangers operate for long periods of time under the same conditions, and they are classified as steady-flow devices. (Note that the flow field near the rotating blades of a turbomachine is of course unsteady, but we consider the overall flow field rather than the details at some localities when we classify
FIGURE 1–18 In this schlieren image of a girl in a swimming suit, the rise of lighter, warmer air adjacent to her body indicates that humans and warmblooded animals are surrounded by thermal plumes of rising warm air. G. S. Settles, Gas Dynamics Lab, Penn State University. Used by permission.
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(a)
(b)
FIGURE 1–19 Oscillating wake of a blunt-based airfoil at Mach number 0.6. Photo (a) is an instantaneous image, while photo (b) is a long-exposure (time-averaged) image. (a) Dyment, A., Flodrops, J. P. & Gryson, P. 1982 in Flow Visualization II, W. Merzkirch, ed., 331–336. Washington: Hemisphere. Used by permission of Arthur Dyment. (b) Dyment, A. & Gryson, P. 1978 in Inst. Mèc. Fluides Lille, No. 78-5. Used by permission of Arthur Dyment.
devices.) During steady flow, the fluid properties can change from point to point within a device, but at any fixed point they remain constant. Therefore, the volume, the mass, and the total energy content of a steady-flow device or flow section remain constant in steady operation. Steady-flow conditions can be closely approximated by devices that are intended for continuous operation such as turbines, pumps, boilers, condensers, and heat exchangers of power plants or refrigeration systems. Some cyclic devices, such as reciprocating engines or compressors, do not satisfy the steady-flow conditions since the flow at the inlets and the exits is pulsating and not steady. However, the fluid properties vary with time in a periodic manner, and the flow through these devices can still be analyzed as a steady-flow process by using time-averaged values for the properties. Some fascinating visualizations of fluid flow are provided in the book An Album of Fluid Motion by Milton Van Dyke (1982). A nice illustration of an unsteady-flow field is shown in Fig. 1–19, taken from Van Dyke’s book. Figure 1–19a is an instantaneous snapshot from a high-speed motion picture; it reveals large, alternating, swirling, turbulent eddies that are shed into the periodically oscillating wake from the blunt base of the object. The eddies produce shock waves that move upstream alternately over the top and bottom surfaces of the airfoil in an unsteady fashion. Figure 1–19b shows the same flow field, but the film is exposed for a longer time so that the image is time averaged over 12 cycles. The resulting time-averaged flow field appears “steady” since the details of the unsteady oscillations have been lost in the long exposure. One of the most important jobs of an engineer is to determine whether it is sufficient to study only the time-averaged “steady” flow features of a problem, or whether a more detailed study of the unsteady features is required. If the engineer were interested only in the overall properties of the flow field, (such as the time-averaged drag coefficient, the mean velocity, and pressure fields) a time-averaged description like that of Fig. 1–19b, time-averaged experimental measurements, or an analytical or numerical calculation of the time-averaged flow field would be sufficient. However, if the engineer were interested in details about the unsteady-flow field, such as flow-induced vibrations, unsteady pressure fluctuations, or the sound waves emitted from the turbulent eddies or the shock waves, a time-averaged description of the flow field would be insufficient. Most of the analytical and computational examples provided in this textbook deal with steady or time-averaged flows, although we occasionally point out some relevant unsteady-flow features as well when appropriate.
One-, Two-, and Three-Dimensional Flows A flow field is best characterized by the velocity distribution, and thus a flow is said to be one-, two-, or three-dimensional if the flow velocity varies in one, two, or three primary dimensions, respectively. A typical fluid flow involves a three-dimensional geometry, and the velocity may vary in all three dimensions, rendering the flow three-dimensional [V (x, y, z) in rectangular or V (r, u, z) in cylindrical coordinates]. However, the variation of velocity in certain directions can be small relative to the variation in other directions and can be ignored with negligible error. In such cases, the flow can be modeled conveniently as being one- or two-dimensional, which is easier to analyze. →
→
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Fully developed velocity profile, V(r)
r
z
FIGURE 1–20 The development of the velocity profile in a circular pipe. V ! V(r, z) and thus the flow is two-dimensional in the entrance region, and becomes one-dimensional downstream when the velocity profile fully develops and remains unchanged in the flow direction, V ! V(r).
Consider steady flow of a fluid through a circular pipe attached to a large tank. The fluid velocity everywhere on the pipe surface is zero because of the no-slip condition, and the flow is two-dimensional in the entrance region of the pipe since the velocity changes in both the r- and z-directions. The velocity profile develops fully and remains unchanged after some distance from the inlet (about 10 pipe diameters in turbulent flow, and less in laminar pipe flow, as in Fig. 1–20), and the flow in this region is said to be fully developed. The fully developed flow in a circular pipe is one-dimensional since the velocity varies in the radial r-direction but not in the angular u- or axial z-directions, as shown in Fig. 1–20. That is, the velocity profile is the same at any axial z-location, and it is symmetric about the axis of the pipe. Note that the dimensionality of the flow also depends on the choice of coordinate system and its orientation. The pipe flow discussed, for example, is one-dimensional in cylindrical coordinates, but two-dimensional in Cartesian coordinates—illustrating the importance of choosing the most appropriate coordinate system. Also note that even in this simple flow, the velocity cannot be uniform across the cross section of the pipe because of the no-slip condition. However, at a well-rounded entrance to the pipe, the velocity profile may be approximated as being nearly uniform across the pipe, since the velocity is nearly constant at all radii except very close to the pipe wall. A flow may be approximated as two-dimensional when the aspect ratio is large and the flow does not change appreciably along the longer dimension. For example, the flow of air over a car antenna can be considered two-dimensional except near its ends since the antenna’s length is much greater than its diameter, and the airflow hitting the antenna is fairly uniform (Fig. 1–21).
EXAMPLE 1–1
FIGURE 1–21 Flow over a car antenna is approximately two-dimensional except near the top and bottom of the antenna.
Axisymmetric Flow over a Bullet
Consider a bullet piercing through calm air. Determine if the time-averaged airflow over the bullet during its flight is one-, two-, or three-dimensional (Fig. 1–22).
SOLUTION It is to be determined whether airflow over a bullet is one-, two-, or three-dimensional. Assumptions There are no significant winds and the bullet is not spinning. Analysis The bullet possesses an axis of symmetry and is therefore an axisymmetric body. The airflow upstream of the bullet is parallel to this axis, and we expect the time-averaged airflow to be rotationally symmetric about
Axis of symmetry r z
u
FIGURE 1–22 Axisymmetric flow over a bullet.
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the axis—such flows are said to be axisymmetric. The velocity in this case varies with axial distance z and radial distance r, but not with angle u. Therefore, the time-averaged airflow over the bullet is two-dimensional. Discussion While the time-averaged airflow is axisymmetric, the instantaneous airflow is not, as illustrated in Fig. 1–19.
1–5 SURROUNDINGS
SYSTEM
BOUNDARY
FIGURE 1–23 System, surroundings, and boundary.
Moving boundary GAS 2 kg 1.5 m3
GAS 2 kg 1 m3
Fixed boundary
FIGURE 1–24 A closed system with a moving boundary.
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SYSTEM AND CONTROL VOLUME
A system is defined as a quantity of matter or a region in space chosen for study. The mass or region outside the system is called the surroundings. The real or imaginary surface that separates the system from its surroundings is called the boundary (Fig. 1–23). The boundary of a system can be fixed or movable. Note that the boundary is the contact surface shared by both the system and the surroundings. Mathematically speaking, the boundary has zero thickness, and thus it can neither contain any mass nor occupy any volume in space. Systems may be considered to be closed or open, depending on whether a fixed mass or a volume in space is chosen for study. A closed system (also known as a control mass) consists of a fixed amount of mass, and no mass can cross its boundary. But energy, in the form of heat or work, can cross the boundary, and the volume of a closed system does not have to be fixed. If, as a special case, even energy is not allowed to cross the boundary, that system is called an isolated system. Consider the piston–cylinder device shown in Fig. 1–24. Let us say that we would like to find out what happens to the enclosed gas when it is heated. Since we are focusing our attention on the gas, it is our system. The inner surfaces of the piston and the cylinder form the boundary, and since no mass is crossing this boundary, it is a closed system. Notice that energy may cross the boundary, and part of the boundary (the inner surface of the piston, in this case) may move. Everything outside the gas, including the piston and the cylinder, is the surroundings. An open system, or a control volume, as it is often called, is a properly selected region in space. It usually encloses a device that involves mass flow such as a compressor, turbine, or nozzle. Flow through these devices is best studied by selecting the region within the device as the control volume. Both mass and energy can cross the boundary of a control volume. A large number of engineering problems involve mass flow in and out of a system and, therefore, are modeled as control volumes. A water heater, a car radiator, a turbine, and a compressor all involve mass flow and should be analyzed as control volumes (open systems) instead of as control masses (closed systems). In general, any arbitrary region in space can be selected as a control volume. There are no concrete rules for the selection of control volumes, but the proper choice certainly makes the analysis much easier. If we were to analyze the flow of air through a nozzle, for example, a good choice for the control volume would be the region within the nozzle. A control volume can be fixed in size and shape, as in the case of a nozzle, or it may involve a moving boundary, as shown in Fig. 1–25. Most control volumes, however, have fixed boundaries and thus do not involve any
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Imaginary boundary
FIGURE 1–25 A control volume may involve fixed, moving, real, and imaginary boundaries.
Real boundary
CV (a nozzle)
Moving boundary CV Fixed boundary
(a) A control volume (CV) with real and imaginary boundaries
(b) A control volume (CV) with fixed and moving boundaries
moving boundaries. A control volume may also involve heat and work interactions just as a closed system, in addition to mass interaction.
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IMPORTANCE OF DIMENSIONS AND UNITS
Any physical quantity can be characterized by dimensions. The magnitudes assigned to the dimensions are called units. Some basic dimensions such as mass m, length L, time t, and temperature T are selected as primary or fundamental dimensions, while others such as velocity V, energy E, and volume V are expressed in terms of the primary dimensions and are called secondary dimensions, or derived dimensions. A number of unit systems have been developed over the years. Despite strong efforts in the scientific and engineering community to unify the world with a single unit system, two sets of units are still in common use today: the English system, which is also known as the United States Customary System (USCS), and the metric SI (from Le Système International d’ Unités), which is also known as the International System. The SI is a simple and logical system based on a decimal relationship between the various units, and it is being used for scientific and engineering work in most of the industrialized nations, including England. The English system, however, has no apparent systematic numerical base, and various units in this system are related to each other rather arbitrarily (12 in ! 1 ft, 1 mile ! 5280 ft, 4 qt ! 1 gal, etc.), which makes it confusing and difficult to learn. The United States is the only industrialized country that has not yet fully converted to the metric system. The systematic efforts to develop a universally acceptable system of units dates back to 1790 when the French National Assembly charged the French Academy of Sciences to come up with such a unit system. An early version of the metric system was soon developed in France, but it did not find universal acceptance until 1875 when The Metric Convention Treaty was prepared and signed by 17 nations, including the United States. In this international treaty, meter and gram were established as the metric units for length and mass, respectively, and a General Conference of Weights and Measures (CGPM) was established that was to meet every six years. In 1960, the
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TA B L E 1 – 1 The seven fundamental (or primary) dimensions and their units in SI Dimension
Unit
Length Mass Time Temperature Electric current Amount of light Amount of matter
meter (m) kilogram (kg) second (s) kelvin (K) ampere (A) candela (cd) mole (mol)
TA B L E 1 – 2 Standard prefixes in SI units Multiple 10 109 106 103 102 101 10$1 10$2 10$3 10$6 10$9 10$12 12
Prefix tera, T giga, G mega, M kilo, k hecto, h deka, da deci, d centi, c milli, m micro, m nano, n pico, p
CGPM produced the SI, which was based on six fundamental quantities, and their units were adopted in 1954 at the Tenth General Conference of Weights and Measures: meter (m) for length, kilogram (kg) for mass, second (s) for time, ampere (A) for electric current, degree Kelvin (°K) for temperature, and candela (cd) for luminous intensity (amount of light). In 1971, the CGPM added a seventh fundamental quantity and unit: mole (mol) for the amount of matter. Based on the notational scheme introduced in 1967, the degree symbol was officially dropped from the absolute temperature unit, and all unit names were to be written without capitalization even if they were derived from proper names (Table 1–1). However, the abbreviation of a unit was to be capitalized if the unit was derived from a proper name. For example, the SI unit of force, which is named after Sir Isaac Newton (1647–1723), is newton (not Newton), and it is abbreviated as N. Also, the full name of a unit may be pluralized, but its abbreviation cannot. For example, the length of an object can be 5 m or 5 meters, not 5 ms or 5 meter. Finally, no period is to be used in unit abbreviations unless they appear at the end of a sentence. For example, the proper abbreviation of meter is m (not m.). The recent move toward the metric system in the United States seems to have started in 1968 when Congress, in response to what was happening in the rest of the world, passed a Metric Study Act. Congress continued to promote a voluntary switch to the metric system by passing the Metric Conversion Act in 1975. A trade bill passed by Congress in 1988 set a September 1992 deadline for all federal agencies to convert to the metric system. However, the deadlines were relaxed later with no clear plans for the future. The industries that are heavily involved in international trade (such as the automotive, soft drink, and liquor industries) have been quick in converting to the metric system for economic reasons (having a single worldwide design, fewer sizes, smaller inventories, etc.). Today, nearly all the cars manufactured in the United States are metric. Most car owners probably do not realize this until they try an English socket wrench on a metric bolt. Most industries, however, resisted the change, thus slowing down the conversion process. Presently the United States is a dual-system society, and it will stay that way until the transition to the metric system is completed. This puts an extra burden on today’s engineering students, since they are expected to retain their understanding of the English system while learning, thinking, and working in terms of the SI. Given the position of the engineers in the transition period, both unit systems are used in this text, with particular emphasis on SI units. As pointed out, the SI is based on a decimal relationship between units. The prefixes used to express the multiples of the various units are listed in Table 1–2. They are standard for all units, and the student is encouraged to memorize them because of their widespread use (Fig. 1–26).
Some SI and English Units In SI, the units of mass, length, and time are the kilogram (kg), meter (m), and second (s), respectively. The respective units in the English system are the pound-mass (lbm), foot (ft), and second (s). The pound symbol lb is
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200 mL (0.2 L)
1 kg (10 3 g)
FIGURE 1–26 The SI unit prefixes are used in all branches of engineering.
1 M# (10 6 #)
actually the abbreviation of libra, which was the ancient Roman unit of weight. The English retained this symbol even after the end of the Roman occupation of Britain in 410. The mass and length units in the two systems are related to each other by 1 lbm ! 0.45359 kg 1 ft ! 0.3048 m
In the English system, force is usually considered to be one of the primary dimensions and is assigned a nonderived unit. This is a source of confusion and error that necessitates the use of a dimensional constant (gc) in many formulas. To avoid this nuisance, we consider force to be a secondary dimension whose unit is derived from Newton’s second law, i.e.,
m = 1 kg
a = 1 m/s 2
Force ! (Mass) (Acceleration)
or
F ! ma
(1–1)
m = 32.174 lbm
In SI, the force unit is the newton (N), and it is defined as the force required to accelerate a mass of 1 kg at a rate of 1 m/s2. In the English system, the force unit is the pound-force (lbf) and is defined as the force required to accelerate a mass of 32.174 lbm (1 slug) at a rate of 1 ft/s2 (Fig. 1–27). That is,
a = 1 ft/s 2
F=1N
F = 1 lbf
FIGURE 1–27 The definition of the force units. 1 kgf
1 N ! 1 kg " m/s2 1 lbf ! 32.174 lbm " ft/s2
A force of 1 N is roughly equivalent to the weight of a small apple (m ! 102 g), whereas a force of 1 lbf is roughly equivalent to the weight of four medium apples (mtotal ! 454 g), as shown in Fig. 1–28. Another force unit in common use in many European countries is the kilogram-force (kgf), which is the weight of 1 kg mass at sea level (1 kgf ! 9.807 N). The term weight is often incorrectly used to express mass, particularly by the “weight watchers.” Unlike mass, weight W is a force. It is the gravitational force applied to a body, and its magnitude is determined from Newton’s second law, W ! mg (N)
10 apples m = 1 kg 1 apple m = 102 g
1N
4 apples m = 1 lbm
1 lbf
(1–2)
where m is the mass of the body, and g is the local gravitational acceleration (g is 9.807 m/s2 or 32.174 ft/s2 at sea level and 45° latitude). An ordinary bathroom scale measures the gravitational force acting on a body. The weight of a unit volume of a substance is called the specific weight g and is determined from g ! rg, where r is density. The mass of a body remains the same regardless of its location in the universe. Its weight, however, changes with a change in gravitational acceleration. A body weighs less on top of a mountain since g decreases with altitude.
FIGURE 1–28 The relative magnitudes of the force units newton (N), kilogram-force (kgf), and pound-force (lbf).
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FIGURE 1–29 A body weighing 150 lbf on earth will weigh only 25 lbf on the moon.
kg g = 9.807 m/s2 W = 9.807 kg · m/s2 = 9.807 N = 1 kgf
lbm
g = 32.174 ft/s2 W = 32.174 lbm · ft/s2 = 1 lbf
FIGURE 1–30 The weight of a unit mass at sea level.
On the surface of the moon, an astronaut weighs about one-sixth of what she or he normally weighs on earth (Fig. 1–29). At sea level a mass of 1 kg weighs 9.807 N, as illustrated in Fig. 1–30. A mass of 1 lbm, however, weighs 1 lbf, which misleads people to believe that pound-mass and pound-force can be used interchangeably as pound (lb), which is a major source of error in the English system. It should be noted that the gravity force acting on a mass is due to the attraction between the masses, and thus it is proportional to the magnitudes of the masses and inversely proportional to the square of the distance between them. Therefore, the gravitational acceleration g at a location depends on the local density of the earth’s crust, the distance to the center of the earth, and to a lesser extent, the positions of the moon and the sun. The value of g varies with location from 9.8295 m/s2 at 4500 m below sea level to 7.3218 m/s2 at 100,000 m above sea level. However, at altitudes up to 30,000 m, the variation of g from the sea-level value of 9.807 m/s2 is less than 1 percent. Therefore, for most practical purposes, the gravitational acceleration can be assumed to be constant at 9.81 m/s2. It is interesting to note that at locations below sea level, the value of g increases with distance from the sea level, reaches a maximum at about 4500 m, and then starts decreasing. (What do you think the value of g is at the center of the earth?) The primary cause of confusion between mass and weight is that mass is usually measured indirectly by measuring the gravity force it exerts. This approach also assumes that the forces exerted by other effects such as air buoyancy and fluid motion are negligible. This is like measuring the distance to a star by measuring its red shift, or measuring the altitude of an airplane by measuring barometric pressure. Both of these are also indirect measurements. The correct direct way of measuring mass is to compare it to a known mass. This is cumbersome, however, and it is mostly used for calibration and measuring precious metals. Work, which is a form of energy, can simply be defined as force times distance; therefore, it has the unit “newton-meter (N . m),” which is called a joule (J). That is, 1J!1N%m
(1–3)
A more common unit for energy in SI is the kilojoule (1 kJ ! 103 J). In the English system, the energy unit is the Btu (British thermal unit), which is defined as the energy required to raise the temperature of 1 lbm of water at 68°F by 1°F. In the metric system, the amount of energy needed to raise the temperature of 1 g of water at 14.5°C by 1°C is defined as 1 calorie (cal), and 1 cal ! 4.1868 J. The magnitudes of the kilojoule and Btu are almost identical (1 Btu ! 1.0551 kJ).
Dimensional Homogeneity FIGURE 1–31 To be dimensionally homogeneous, all the terms in an equation must have the same unit. © Reprinted with special permission of King Features Syndicate.
We all know from grade school that apples and oranges do not add. But we somehow manage to do it (by mistake, of course). In engineering, all equations must be dimensionally homogeneous. That is, every term in an equation must have the same unit (Fig. 1–31). If, at some stage of an analysis, we find ourselves in a position to add two quantities that have different units, it is a clear indication that we have made an error at an earlier stage. So checking dimensions can serve as a valuable tool to spot errors.
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EXAMPLE 1–2
Spotting Errors from Unit Inconsistencies
While solving a problem, a person ended up with the following equation at some stage:
E ! 25 kJ & 7 kJ/kg where E is the total energy and has the unit of kilojoules. Determine how to correct the error and discuss what may have caused it. SOLUTION During an analysis, a relation with inconsistent units is obtained. A correction is to be found, and the probable cause of the error is to be determined. Analysis The two terms on the right-hand side do not have the same units, and therefore they cannot be added to obtain the total energy. Multiplying the last term by mass will eliminate the kilograms in the denominator, and the whole equation will become dimensionally homogeneous; that is, every term in the equation will have the same unit. Discussion Obviously this error was caused by forgetting to multiply the last term by mass at an earlier stage.
We all know from experience that units can give terrible headaches if they are not used carefully in solving a problem. However, with some attention and skill, units can be used to our advantage. They can be used to check formulas; they can even be used to derive formulas, as explained in the following example. EXAMPLE 1–3
Obtaining Formulas from Unit Considerations
A tank is filled with oil whose density is r ! 850 kg/m3. If the volume of the tank is V ! 2 m3, determine the amount of mass m in the tank.
SOLUTION The volume of an oil tank is given. The mass of oil is to be determined. Assumptions Oil is an incompressible substance and thus its density is constant. Analysis A sketch of the system just described is given in Fig. 1–32. Suppose we forgot the formula that relates mass to density and volume. However, we know that mass has the unit of kilograms. That is, whatever calculations we do, we should end up with the unit of kilograms. Putting the given information into perspective, we have r ! 850 kg/m
3
and
V!2m
3
It is obvious that we can eliminate m and end up with kg by multiplying these two quantities. Therefore, the formula we are looking for should be 3
m ! rV Thus,
m ! (850 kg/m3)(2 m3) ! 1700 kg Discussion formulas.
Note that this approach may not work for more complicated
OIL
V = 2 m3 ρ = 850 kg/m3 m=?
FIGURE 1–32 Schematic for Example 1–3.
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The student should keep in mind that a formula that is not dimensionally homogeneous is definitely wrong, but a dimensionally homogeneous formula is not necessarily right.
Unity Conversion Ratios Just as all nonprimary dimensions can be formed by suitable combinations of primary dimensions, all nonprimary units (secondary units) can be formed by combinations of primary units. Force units, for example, can be expressed as N ! kg
m s2
and lbf ! 32.174 lbm
ft s2
They can also be expressed more conveniently as unity conversion ratios as N !1 kg % m/s2
and
lbf !1 32.174 lbm % ft/s2
Unity conversion ratios are identically equal to 1 and are unitless, and thus such ratios (or their inverses) can be inserted conveniently into any calculation to properly convert units. Students are encouraged to always use unity conversion ratios such as those given here when converting units. Some textbooks insert the archaic gravitational constant gc defined as gc ! 32.174 lbm · ft/lbf · s2 ! kg · m/N · s2 ! 1 into equations in order to force units to match. This practice leads to unnecessary confusion and is strongly discouraged by the present authors. We recommend that students instead use unity conversion ratios. EXAMPLE 1–4
The Weight of One Pound-Mass
Using unity conversion ratios, show that 1.00 lbm weighs 1.00 lbf on earth (Fig. 1–33). lbm
FIGURE 1–33 A mass of 1 lbm weighs 1 lbf on earth.
Solution
A mass of 1.00 lbm is subjected to standard earth gravity. Its weight in lbf is to be determined. Assumptions Standard sea-level conditions are assumed. Properties The gravitational constant is g ! 32.174 ft/s2. Analysis We apply Newton’s second law to calculate the weight (force) that corresponds to the known mass and acceleration. The weight of any object is equal to its mass times the local value of gravitational acceleration. Thus,
1 lbf W ! mg ! (1.00 lbm)(32.174 ft/s2)a b ! 1.00 lbf 32.174 lbm % ft/s2
Discussion Mass is the same regardless of its location. However, on some other planet with a different value of gravitational acceleration, the weight of 1 lbm would differ from that calculated here.
When you buy a box of breakfast cereal, the printing may say “Net weight: One pound (454 grams).” (See Fig. 1–34.) Technically, this means that the cereal inside the box weighs 1.00 lbf on earth and has a mass of
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453.6 gm (0.4536 kg). Using Newton’s second law, the actual weight on earth of the cereal in the metric system is W ! mg ! (453.6 g)(9.81 m/s2) a
1–7
■
1 kg 1N ba b ! 4.49 N 2 1000 g 1 kg % m/s
Net weight: One pound (454 grams)
MATHEMATICAL MODELING OF ENGINEERING PROBLEMS
An engineering device or process can be studied either experimentally (testing and taking measurements) or analytically (by analysis or calculations). The experimental approach has the advantage that we deal with the actual physical system, and the desired quantity is determined by measurement, within the limits of experimental error. However, this approach is expensive, time-consuming, and often impractical. Besides, the system we are studying may not even exist. For example, the entire heating and plumbing systems of a building must usually be sized before the building is actually built on the basis of the specifications given. The analytical approach (including the numerical approach) has the advantage that it is fast and inexpensive, but the results obtained are subject to the accuracy of the assumptions, approximations, and idealizations made in the analysis. In engineering studies, often a good compromise is reached by reducing the choices to just a few by analysis, and then verifying the findings experimentally.
FIGURE 1–34 A quirk in the metric system of units.
Modeling in Engineering The descriptions of most scientific problems involve equations that relate the changes in some key variables to each other. Usually the smaller the increment chosen in the changing variables, the more general and accurate the description. In the limiting case of infinitesimal or differential changes in variables, we obtain differential equations that provide precise mathematical formulations for the physical principles and laws by representing the rates of change as derivatives. Therefore, differential equations are used to investigate a wide variety of problems in sciences and engineering (Fig. 1–35). However, many problems encountered in practice can be solved without resorting to differential equations and the complications associated with them. The study of physical phenomena involves two important steps. In the first step, all the variables that affect the phenomena are identified, reasonable assumptions and approximations are made, and the interdependence of these variables is studied. The relevant physical laws and principles are invoked, and the problem is formulated mathematically. The equation itself is very instructive as it shows the degree of dependence of some variables on others, and the relative importance of various terms. In the second step, the problem is solved using an appropriate approach, and the results are interpreted. Many processes that seem to occur in nature randomly and without any order are, in fact, being governed by some visible or not-so-visible physical laws. Whether we notice them or not, these laws are there, governing consistently and predictably over what seem to be ordinary events. Most of
Physical problem Identify important variables
Apply relevant physical laws
Make reasonable assumptions and approximations
A differential equation Apply applicable solution technique
Apply boundary and initial conditions
Solution of the problem
FIGURE 1–35 Mathematical modeling of physical problems.
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AY
SY
W
EA
HARD WAY
SOLUTION
PROBLEM
FIGURE 1–36 A step-by-step approach can greatly simplify problem solving.
these laws are well defined and well understood by scientists. This makes it possible to predict the course of an event before it actually occurs or to study various aspects of an event mathematically without actually running expensive and time-consuming experiments. This is where the power of analysis lies. Very accurate results to meaningful practical problems can be obtained with relatively little effort by using a suitable and realistic mathematical model. The preparation of such models requires an adequate knowledge of the natural phenomena involved and the relevant laws, as well as sound judgment. An unrealistic model will obviously give inaccurate and thus unacceptable results. An analyst working on an engineering problem often finds himself or herself in a position to make a choice between a very accurate but complex model, and a simple but not-so-accurate model. The right choice depends on the situation at hand. The right choice is usually the simplest model that yields satisfactory results. Also, it is important to consider the actual operating conditions when selecting equipment. Preparing very accurate but complex models is usually not so difficult. But such models are not much use to an analyst if they are very difficult and time-consuming to solve. At the minimum, the model should reflect the essential features of the physical problem it represents. There are many significant real-world problems that can be analyzed with a simple model. But it should always be kept in mind that the results obtained from an analysis are at best as accurate as the assumptions made in simplifying the problem. Therefore, the solution obtained should not be applied to situations for which the original assumptions do not hold. A solution that is not quite consistent with the observed nature of the problem indicates that the mathematical model used is too crude. In that case, a more realistic model should be prepared by eliminating one or more of the questionable assumptions. This will result in a more complex problem that, of course, is more difficult to solve. Thus any solution to a problem should be interpreted within the context of its formulation.
1–8
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PROBLEM-SOLVING TECHNIQUE
The first step in learning any science is to grasp the fundamentals and to gain a sound knowledge of it. The next step is to master the fundamentals by testing this knowledge. This is done by solving significant real-world problems. Solving such problems, especially complicated ones, requires a systematic approach. By using a step-by-step approach, an engineer can reduce the solution of a complicated problem into the solution of a series of simple problems (Fig. 1–36). When you are solving a problem, we recommend that you use the following steps zealously as applicable. This will help you avoid some of the common pitfalls associated with problem solving.
Step 1: Problem Statement In your own words, briefly state the problem, the key information given, and the quantities to be found. This is to make sure that you understand the problem and the objectives before you attempt to solve the problem.
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Step 2: Schematic Draw a realistic sketch of the physical system involved, and list the relevant information on the figure. The sketch does not have to be something elaborate, but it should resemble the actual system and show the key features. Indicate any energy and mass interactions with the surroundings. Listing the given information on the sketch helps one to see the entire problem at once. Also, check for properties that remain constant during a process (such as temperature during an isothermal process), and indicate them on the sketch.
Given: Air temperature in Denver To be found: Density of air Missing information: Atmospheric pressure Assumption #1: Take P = 1 atm (Inappropriate. Ignores effect of altitude. Will cause more than 15% error.)
Step 3: Assumptions and Approximations State any appropriate assumptions and approximations made to simplify the problem to make it possible to obtain a solution. Justify the questionable assumptions. Assume reasonable values for missing quantities that are necessary. For example, in the absence of specific data for atmospheric pressure, it can be taken to be 1 atm. However, it should be noted in the analysis that the atmospheric pressure decreases with increasing elevation. For example, it drops to 0.83 atm in Denver (elevation 1610 m) (Fig. 1–37).
Step 4: Physical Laws Apply all the relevant basic physical laws and principles (such as the conservation of mass), and reduce them to their simplest form by utilizing the assumptions made. However, the region to which a physical law is applied must be clearly identified first. For example, the increase in speed of water flowing through a nozzle is analyzed by applying conservation of mass between the inlet and outlet of the nozzle.
Assumption #2: Take P = 0.83 atm (Appropriate. Ignores only minor effects such as weather.)
FIGURE 1–37 The assumptions made while solving an engineering problem must be reasonable and justifiable.
Step 5: Properties Determine the unknown properties at known states necessary to solve the problem from property relations or tables. List the properties separately, and indicate their source, if applicable.
Before streamlining
V
Step 6: Calculations Substitute the known quantities into the simplified relations and perform the calculations to determine the unknowns. Pay particular attention to the units and unit cancellations, and remember that a dimensional quantity without a unit is meaningless. Also, don’t give a false implication of high precision by copying all the digits from the screen of the calculator—round the results to an appropriate number of significant digits (Section 1–10).
FD
V
Unreasonable!
After streamlining
Step 7: Reasoning, Verification, and Discussion Check to make sure that the results obtained are reasonable and intuitive, and verify the validity of the questionable assumptions. Repeat the calculations that resulted in unreasonable values. For example, under the same test conditions the aerodynamic drag acting on a car should not increase after streamlining the shape of the car (Fig. 1–38). Also, point out the significance of the results, and discuss their implications. State the conclusions that can be drawn from the results, and any recommendations that can be made from them. Emphasize the limitations
FD
FIGURE 1–38 The results obtained from an engineering analysis must be checked for reasonableness.
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under which the results are applicable, and caution against any possible misunderstandings and using the results in situations where the underlying assumptions do not apply. For example, if you determined that using a larger-diameter pipe in a proposed pipeline will cost an additional $5000 in materials, but it will reduce the annual pumping costs by $3000, indicate that the larger-diameter pipeline will pay for its cost differential from the electricity it saves in less than two years. However, also state that only additional material costs associated with the larger-diameter pipeline are considered in the analysis. Keep in mind that the solutions you present to your instructors, and any engineering analysis presented to others, is a form of communication. Therefore neatness, organization, completeness, and visual appearance are of utmost importance for maximum effectiveness. Besides, neatness also serves as a great checking tool since it is very easy to spot errors and inconsistencies in neat work. Carelessness and skipping steps to save time often end up costing more time and unnecessary anxiety. The approach described here is used in the solved example problems without explicitly stating each step, as well as in the Solutions Manual of this text. For some problems, some of the steps may not be applicable or necessary. For example, often it is not practical to list the properties separately. However, we cannot overemphasize the importance of a logical and orderly approach to problem solving. Most difficulties encountered while solving a problem are not due to a lack of knowledge; rather, they are due to a lack of organization. You are strongly encouraged to follow these steps in problem solving until you develop your own approach that works best for you.
1–9
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ENGINEERING SOFTWARE PACKAGES
You may be wondering why we are about to undertake an in-depth study of the fundamentals of another engineering science. After all, almost all such problems we are likely to encounter in practice can be solved using one of several sophisticated software packages readily available in the market today. These software packages not only give the desired numerical results, but also supply the outputs in colorful graphical form for impressive presentations. It is unthinkable to practice engineering today without using some of these packages. This tremendous computing power available to us at the touch of a button is both a blessing and a curse. It certainly enables engineers to solve problems easily and quickly, but it also opens the door for abuses and misinformation. In the hands of poorly educated people, these software packages are as dangerous as sophisticated powerful weapons in the hands of poorly trained soldiers. Thinking that a person who can use the engineering software packages without proper training on fundamentals can practice engineering is like thinking that a person who can use a wrench can work as a car mechanic. If it were true that the engineering students do not need all these fundamental courses they are taking because practically everything can be done by computers quickly and easily, then it would also be true that the employers would no longer need high-salaried engineers since any person who knows how to use a word-processing program can also learn how to use those software packages. However, the statistics show that the need for engineers is on the rise, not on the decline, despite the availability of these powerful packages.
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We should always remember that all the computing power and the engineering software packages available today are just tools, and tools have meaning only in the hands of masters. Having the best word-processing program does not make a person a good writer, but it certainly makes the job of a good writer much easier and makes the writer more productive (Fig. 1–39). Hand calculators did not eliminate the need to teach our children how to add or subtract, and the sophisticated medical software packages did not take the place of medical school training. Neither will engineering software packages replace the traditional engineering education. They will simply cause a shift in emphasis in the courses from mathematics to physics. That is, more time will be spent in the classroom discussing the physical aspects of the problems in greater detail, and less time on the mechanics of solution procedures. All these marvelous and powerful tools available today put an extra burden on today’s engineers. They must still have a thorough understanding of the fundamentals, develop a “feel” of the physical phenomena, be able to put the data into proper perspective, and make sound engineering judgments, just like their predecessors. However, they must do it much better, and much faster, using more realistic models because of the powerful tools available today. The engineers in the past had to rely on hand calculations, slide rules, and later hand calculators and computers. Today they rely on software packages. The easy access to such power and the possibility of a simple misunderstanding or misinterpretation causing great damage make it more important today than ever to have solid training in the fundamentals of engineering. In this text we make an extra effort to put the emphasis on developing an intuitive and physical understanding of natural phenomena instead of on the mathematical details of solution procedures.
Engineering Equation Solver (EES) EES is a program that solves systems of linear or nonlinear algebraic or differential equations numerically. It has a large library of built-in thermodynamic property functions as well as mathematical functions, and allows the user to supply additional property data. Unlike some software packages, EES does not solve engineering problems; it only solves the equations supplied by the user. Therefore, the user must understand the problem and formulate it by applying any relevant physical laws and relations. EES saves the user considerable time and effort by simply solving the resulting mathematical equations. This makes it possible to attempt significant engineering problems not suitable for hand calculations and to conduct parametric studies quickly and conveniently. EES is a very powerful yet intuitive program that is very easy to use, as shown in Example 1–5. The use and capabilities of EES are explained in Appendix 3 on the enclosed DVD. EXAMPLE 1–5
Solving a System of Equations with EES
The difference of two numbers is 4, and the sum of the squares of these two numbers is equal to the sum of the numbers plus 20. Determine these two numbers.
Attached is a pdf of the text with windows and approx sizes for the art. I'll give you rough ideas on the art, though you may have some different thoughts on approaching these. Fig 1 - 41 x 30 The boxes fall into 2 columns, Type 1/2 on left and Type 1 on right. Nonenzymatic glycation is in the middle, between columns. Oxidative Stress and Axonal Degeneration are common outcomes and should be centered at the bottom beneath both columns (no need to stack them as shown). I wish I knew what the Polyol Pathway was, cause I'd like to illustrate it somehow. Fig 2 -- 41 x 26 This one's kinda straighforward, though I'd push Type 1/2 and Hyperglycemia further to the left, so that everything falls roughly under the other, Type 1 column.
FIGURE 1–39 An excellent word-processing program does not make a person a good writer; it simply makes a good writer a more efficient writer.
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SOLUTION Relations are given for the difference and the sum of the squares of two numbers. They are to be determined. Analysis We start the EES program by double-clicking on its icon, open a new file, and type the following on the blank screen that appears: x–y!4 xˆ2&yˆ2!x&y&20 which is an exact mathematical expression of the problem statement with x and y denoting the unknown numbers. The solution to this system of two nonlinear equations with two unknowns is obtained by a single click on the “calculator” icon on the taskbar. It gives
x!5
and
y!1
Discussion Note that all we did is formulate the problem as we would on paper; EES took care of all the mathematical details of solution. Also note that equations can be linear or nonlinear, and they can be entered in any order with unknowns on either side. Friendly equation solvers such as EES allow the user to concentrate on the physics of the problem without worrying about the mathematical complexities associated with the solution of the resulting system of equations.
FLUENT FLUENT is a computational fluid dynamics (CFD) code widely used for flow-modeling applications. The first step in analysis is preprocessing, which involves building a model or importing one from a CAD package, applying a finite-volume-based mesh, and entering data. Once the numerical model is prepared, FLUENT performs the necessary calculations and produces the desired results. The final step in analysis is postprocessing, which involves organization and interpretation of the data and images. Packages tailored for specific applications such as electronics cooling, ventilating systems, and mixing are also available. FLUENT can handle subsonic or supersonic flows, steady or transient flows, laminar or turbulent flows, Newtonian or non-Newtonian flows, single or multiphase flows, chemical reactions including combustion, flow through porous media, heat transfer, and flowinduced vibrations. Most numerical solutions presented in this text are obtained using FLUENT, and CFD is discussed in more detail in Chap. 15.
1–10
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ACCURACY, PRECISION, AND SIGNIFICANT DIGITS
In engineering calculations, the supplied information is not known to more than a certain number of significant digits, usually three digits. Consequently, the results obtained cannot possibly be precise to more significant digits. Reporting results in more significant digits implies greater precision than exists, and it should be avoided. Regardless of the system of units employed, engineers must be aware of three principles that govern the proper use of numbers: accuracy, precision,
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and significant digits. For engineering measurements, they are defined as follows:
+++++ ++ +
• Accuracy error (inaccuracy) is the value of one reading minus the true value. In general, accuracy of a set of measurements refers to the closeness of the average reading to the true value. Accuracy is generally associated with repeatable, fixed errors. • Precision error is the value of one reading minus the average of readings. In general, precision of a set of measurements refers to the fineness of the resolution and the repeatability of the instrument. Precision is generally associated with unrepeatable, random errors. • Significant digits are digits that are relevant and meaningful.
A
A measurement or calculation can be very precise without being very accurate, and vice versa. For example, suppose the true value of wind speed is 25.00 m/s. Two anemometers A and B take five wind speed readings each:
+
+
Anemometer A: 25.50, 25.69, 25.52, 25.58, and 25.61 m/s. Average of all readings ! 25.58 m/s. Anemometer B: 26.3, 24.5, 23.9, 26.8, and 23.6 m/s. Average of all readings ! 25.02 m/s.
Clearly, anemometer A is more precise, since none of the readings differs by more than 0.11 m/s from the average. However, the average is 25.58 m/s, 0.58 m/s greater than the true wind speed; this indicates significant bias error, also called constant error or systematic error. On the other hand, anemometer B is not very precise, since its readings swing wildly from the average; but its overall average is much closer to the true value. Hence, anemometer B is more accurate than anemometer A, at least for this set of readings, even though it is less precise. The difference between accuracy and precision can be illustrated effectively by analogy to shooting a gun at a target, as sketched in Fig. 1–40. Shooter A is very precise, but not very accurate, while shooter B has better overall accuracy, but less precision. Many engineers do not pay proper attention to the number of significant digits in their calculations. The least significant numeral in a number implies the precision of the measurement or calculation. For example, a result written as 1.23 (three significant digits) implies that the result is precise to within one digit in the second decimal place; i.e., the number is somewhere between 1.22 and 1.24. Expressing this number with any more digits would be misleading. The number of significant digits is most easily evaluated when the number is written in exponential notation; the number of significant digits can then simply be counted, including zeroes. Some examples are shown in Table 1–3. When performing calculations or manipulations of several parameters, the final result is generally only as precise as the least precise parameter in the problem. For example, suppose A and B are multiplied to obtain C. If A ! 2.3601 (five significant digits), and B ! 0.34 (two significant digits), then C ! 0.80 (only two digits are significant in the final result). Note that most students are tempted to write C ! 0.802434, with six significant digits, since that is what is displayed on a calculator after multiplying these two numbers.
+ +
+ +
+
B
FIGURE 1–40 Illustration of accuracy versus precision. Shooter A is more precise, but less accurate, while shooter B is more accurate, but less precise.
TA B L E 1 – 3 Significant digits
Number
Number of Exponential Significant Notation Digits
12.3 1.23 ' 101 123,000 1.23 ' 105 0.00123 1.23 ' 10$3 40,300 4.03 ' 104 40,300. 4.0300 ' 104 0.005600 5.600 ' 10$3 0.0056 5.6 ' 10$3 0.006 6. ' 10$3
3 3 3 3 5 4 2 1
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28 FLUID MECHANICS
Given: Volume: V = 3.75 L Density: r = 0.845 kg/L (3 significant digits) Also, 3.75 × 0.845 = 3.16875 Find: Mass: m = rV = 3.16875 kg Rounding to 3 significant digits: m = 3.17 kg
FIGURE 1–41 A result with more significant digits than that of given data falsely implies more precision.
Let’s analyze this simple example carefully. Suppose the exact value of B is 0.33501, which is read by the instrument as 0.34. Also suppose A is exactly 2.3601, as measured by a more accurate and precise instrument. In this case, C ! A ' B ! 0.79066 to five significant digits. Note that our first answer, C ! 0.80 is off by one digit in the second decimal place. Likewise, if B is 0.34499, and is read by the instrument as 0.34, the product of A and B would be 0.81421 to five significant digits. Our original answer of 0.80 is again off by one digit in the second decimal place. The main point here is that 0.80 (to two significant digits) is the best one can expect from this multiplication since, to begin with, one of the values had only two significant digits. Another way of looking at this is to say that beyond the first two digits in the answer, the rest of the digits are meaningless or not significant. For example, if one reports what the calculator displays, 2.3601 times 0.34 equals 0.802434, the last four digits are meaningless. As shown, the final result may lie between 0.79 and 0.81—any digits beyond the two significant digits are not only meaningless, but misleading, since they imply to the reader more precision than is really there. As another example, consider a 3.75-L container filled with gasoline whose density is 0.845 kg/L, and determine its mass. Probably the first thought that comes to your mind is to multiply the volume and density to obtain 3.16875 kg for the mass, which falsely implies that the mass so determined is precise to six significant digits. In reality, however, the mass cannot be more precise than three significant digits since both the volume and the density are precise to three significant digits only. Therefore, the result should be rounded to three significant digits, and the mass should be reported to be 3.17 kg instead of what the calculator displays (Fig. 1–41). The result 3.16875 kg would be correct only if the volume and density were given to be 3.75000 L and 0.845000 kg/L, respectively. The value 3.75 L implies that we are fairly confident that the volume is precise within (0.01 L, and it cannot be 3.74 or 3.76 L. However, the volume can be 3.746, 3.750, 3.753, etc., since they all round to 3.75 L. You should also be aware that sometimes we knowingly introduce small errors in order to avoid the trouble of searching for more accurate data. For example, when dealing with liquid water, we often use the value of 1000 kg/m3 for density, which is the density value of pure water at 0°C. Using this value at 75°C will result in an error of 2.5 percent since the density at this temperature is 975 kg/m3. The minerals and impurities in the water will introduce additional error. This being the case, you should have no reservation in rounding the final results to a reasonable number of significant digits. Besides, having a few percent uncertainty in the results of engineering analysis is usually the norm, not the exception. When writing intermediate results in a computation, it is advisable to keep several “extra” digits to avoid round-off errors; however, the final result should be written with the number of significant digits taken into consideration. The reader must also keep in mind that a certain number of significant digits of precision in the result does not necessarily imply the same number of digits of overall accuracy. Bias error in one of the readings may, for example, significantly reduce the overall accuracy of the result, perhaps even rendering the last significant digit meaningless, and reducing the overall number of reliable digits by one. Experimentally determined values are
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29 CHAPTER 1
subject to measurement errors, and such errors are reflected in the results obtained. For example, if the density of a substance has an uncertainty of 2 percent, then the mass determined using this density value will also have an uncertainty of 2 percent. Finally, when the number of significant digits is unknown, the accepted engineering standard is three significant digits. Therefore, if the length of a pipe is given to be 40 m, we will assume it to be 40.0 m in order to justify using three significant digits in the final results.
EXAMPLE 1–6
Significant Digits and Volume Flow Rate
Jennifer is conducting an experiment that uses cooling water from a garden hose. In order to calculate the volume flow rate of water through the hose, she times how long it takes to fill a container (Fig. 1–42). The volume of water collected is V ! 1.1 gal in time period )t ! 45.62 s, as measured with a stopwatch. Calculate the volume flow rate of water through the hose in units of cubic meters per minute.
SOLUTION Volume flow rate is to be determined from measurements of volume and time period. Assumptions 1 Jennifer recorded her measurements properly, such that the volume measurement is precise to two significant digits while the time period is precise to four significant digits. 2 No water is lost due to splashing out of the container. . Analysis Volume flow rate V is volume displaced per unit time and is expressed as Volume flow rate:
# )V V! )t
Substituting the measured values, the volume flow rate is determined to be
1.1 gal 3.785 ' 10 $3 m3 # 60 s V! a b a b ! 5.5 " 10 #3 m3/min 45.62 s 1 gal 1 min
Discussion The final result is listed to two significant digits since we cannot be confident of any more precision than that. If this were an intermediate step in subsequent calculations, a few extra digits would be carried along to avoid accumulated. round-off error. In such a case, the volume flow rate would be written as V ! 5.4759 ' 10$3 m3/min. Based on the given information, we cannot say anything about the accuracy of our result, since we have no information about systematic errors in either the volume measurement or the time measurement. Also keep in mind that good precision does not guarantee good accuracy. For example, if the batteries in the stopwatch were weak, its accuracy could be quite poor, yet the readout would still be displayed to four significant digits of precision. In common practice, precision is often associated with resolution, which is a measure of how finely the instrument can report the measurement. For example, a digital voltmeter with five digits on its display is said to be more
Hose
Container
FIGURE 1–42 Schematic for Example 1–6 for the measurement of volume flow rate.
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30 FLUID MECHANICS
FIGURE 1–43 An instrument with many digits of resolution (stopwatch c) may be less accurate than an instrument with few digits of resolution (stopwatch a). What can you say about stopwatches b and d?
Exact time span = 45.623451 . . . s
TIMEXAM
TIMEXAM
TIMEXAM
46.
43.
44.189
s
(a)
(b)
s
(c)
TIMEXAM
45.624 s
s
(d)
precise than a digital voltmeter with only three digits. However, the number of displayed digits has nothing to do with the overall accuracy of the measurement. An instrument can be very precise without being very accurate when there are significant bias errors. Likewise, an instrument with very few displayed digits can be more accurate than one with many digits (Fig. 1–43).
SUMMARY In this chapter some basic concepts of fluid mechanics are introduced and discussed. A substance in the liquid or gas phase is referred to as a fluid. Fluid mechanics is the science that deals with the behavior of fluids at rest or in motion and the interaction of fluids with solids or other fluids at the boundaries. The flow of an unbounded fluid over a surface is external flow, and the flow in a pipe or duct is internal flow if the fluid is completely bounded by solid surfaces. A fluid flow is classified as being compressible or incompressible, depending on the density variation of the fluid during flow. The densities of liquids are essentially constant, and thus the flow of liquids is typically incompressible. The term steady implies no change with time. The opposite of steady is unsteady, or transient. The term uniform implies no change with location over a specified region. A flow is said to be one-dimensional when the velocity changes in one dimension only. A fluid in direct contact with a solid surface sticks to the surface and
there is no slip. This is known as the no-slip condition, which leads to the formation of boundary layers along solid surfaces. A system of fixed mass is called a closed system, and a system that involves mass transfer across its boundaries is called an open system or control volume. A large number of engineering problems involve mass flow in and out of a system and are therefore modeled as control volumes. In engineering calculations, it is important to pay particular attention to the units of the quantities to avoid errors caused by inconsistent units, and to follow a systematic approach. It is also important to recognize that the information given is not known to more than a certain number of significant digits, and the results obtained cannot possibly be accurate to more significant digits. The information given on dimensions and units; problem-solving technique; and accuracy, precision, and significant digits will be used throughout the entire text.
REFERENCES AND SUGGESTED READING 1. American Society for Testing and Materials. Standards for Metric Practice. ASTM E 380-79, January 1980. 2. C. T. Crowe, J. A. Roberson, and D. F. Elger. Engineering Fluid Mechanics, 7th ed. New York: Wiley, 2001.
4. G. M. Homsy, H. Aref, K. S. Breuer, S. Hochgreb, J. R. Koseff, B. R. Munson, K. G. Powell, C. R. Robertson, and S. T. Thoroddsen. Multi-Media Fluid Mechanics (CD). Cambridge: Cambridge University Press, 2000.
3. R. W. Fox and A. T. McDonald. Introduction to Fluid Mechanics, 5th ed. New York: Wiley, 1999.
5. M. Van Dyke. An Album of Fluid Motion. Stanford, CA: The Parabolic Press, 1982.
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31 CHAPTER 1
APPLICATION SPOTLIGHT
■
What Nuclear Blasts and Raindrops Have in Common
Guest Author: Lorenz Sigurdson, Vortex Fluid Dynamics Lab, University of Alberta Why do the two images in Fig. 1–44 look alike? Figure 1–44b shows an above-ground nuclear test performed by the U.S. Department of Energy in 1957. An atomic blast created a fireball on the order of 100 m in diameter. Expansion is so quick that a compressible flow feature occurs: an expanding spherical shock wave. The image shown in Fig. 1–44a is an everyday innocuous event: an inverted image of a dye-stained water drop after it has fallen into a pool of water, looking from below the pool surface. It could have fallen from your spoon into a cup of coffee, or been a secondary splash after a raindrop hit a lake. Why is there such a strong similarity between these two vastly different events? The application of fundamental principles of fluid mechanics learned in this book will help you understand much of the answer, although one can go much deeper. The water has higher density (Chap. 2) than air, so the drop has experienced negative buoyancy (Chap. 3) as it has fallen through the air before impact. The fireball of hot gas is less dense than the cool air surrounding it, so it has positive buoyancy and rises. The shock wave (Chap. 12) reflecting from the ground also imparts a positive upward force to the fireball. The primary structure at the top of each image is called a vortex ring. This ring is a mini-tornado of concentrated vorticity (Chap. 4) with the ends of the tornado looping around to close on itself. The laws of kinematics (Chap. 4) tell us that this vortex ring will carry the fluid in a direction toward the top of the page. This is expected in both cases from the forces applied and the law of conservation of momentum applied through a control volume analysis (Chap. 5). One could also analyze this problem with differential analysis (Chaps. 9 and 10) or with computational fluid dynamics (Chap. 15). But why does the shape of the tracer material look so similar? This occurs if there is approximate geometric and kinematic similarity (Chap. 7), and if the flow visualization (Chap. 4) technique is similar. The passive tracers of heat and dust for the bomb, and fluorescent dye for the drop, were introduced in a similar manner as noted in the figure caption. Further knowledge of kinematics and vortex dynamics can help explain the similarity of the vortex structure in the images to much greater detail, as discussed by Sigurdson (1997) and Peck and Sigurdson (1994). Look at the lobes dangling beneath the primary vortex ring, the striations in the “stalk,” and the ring at the base of each structure. There is also topological similarity of this structure to other vortex structures occurring in turbulence. Comparison of the drop and bomb has given us a better understanding of how turbulent structures are created and evolve. What other secrets of fluid mechanics are left to be revealed in explaining the similarity between these two flows? References Peck, B., and Sigurdson, L.W., “The Three-Dimensional Vortex Structure of an Impacting Water Drop,” Phys. Fluids, 6(2) (Part 1), p. 564, 1994. Peck, B., Sigurdson, L.W., Faulkner, B., and Buttar, I., “An Apparatus to Study Drop-Formed Vortex Rings,” Meas. Sci. Tech., 6, p. 1538, 1995. Sigurdson, L.W., “Flow Visualization in Turbulent Large-Scale Structure Research,” Chapter 6 in Atlas of Visualization, Vol. III, Flow Visualization Society of Japan, eds., CRC Press, pp. 99–113, 1997.
(a)
(b)
FIGURE 1–44 Comparison of the vortex structure created by: (a) a water drop after impacting a pool of water (inverted, from Peck and Sigurdson, 1994), and (b) an above-ground nuclear test in Nevada in 1957 (U.S. Department of Energy). The 2.6 mm drop was dyed with fluorescent tracer and illuminated by a strobe flash 50 ms after it had fallen 35 mm and impacted the clear pool. The drop was approximately spherical at the time of impact with the clear pool of water. Interruption of a laser beam by the falling drop was used to trigger a timer that controlled the time of the strobe flash after impact of the drop. Details of the careful experimental procedure necessary to create the drop photograph are given by Peck and Sigurdson (1994) and Peck et al. (1995). The tracers added to the flow in the bomb case were primarily heat and dust. The heat is from the original fireball which for this particular test (the “Priscilla” event of Operation Plumbob) was large enough to reach the ground from where the bomb was initially suspended. Therefore, the tracer’s initial geometric condition was a sphere intersecting the ground. (a) From Peck, B., and Sigurdson, L. W., Phys. Fluids, 6(2)(Part 1), 564, 1994. Used by permission of the author. (b) United States Department of Energy. Photo from Lorenz Sigurdson.
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32 FLUID MECHANICS
PROBLEMS* Introduction, Classification, and System 1–1C
Define internal, external, and open-channel flows.
1–2C Define incompressible flow and incompressible fluid. Must the flow of a compressible fluid necessarily be treated as compressible? 1–3C
What is the no-slip condition? What causes it?
1–4C What is forced flow? How does it differ from natural flow? Is flow caused by winds forced or natural flow? 1–5C What is a boundary layer? What causes a boundary layer to develop? 1–6C What is the difference between the classical and the statistical approaches? 1–7C
What is a steady-flow process?
1–8C
Define stress, normal stress, shear stress, and pressure.
1–9C
What are system, surroundings, and boundary?
1–10C When is a system a closed system, and when is it a control volume?
Mass, Force, and Units 1–11C What is the difference between pound-mass and pound-force? 1–12C force?
What is the difference between kg-mass and kg-
1–13C What is the net force acting on a car cruising at a constant velocity of 70 km/h (a) on a level road and (b) on an uphill road? 1–14 A 3-kg plastic tank that has a volume of 0.2 m3 is filled with liquid water. Assuming the density of water is 1000 kg/m3, determine the weight of the combined system. 1–15 Determine the mass and the weight of the air contained in a room whose dimensions are 6 m ' 6 m ' 8 m. Assume the density of the air is 1.16 kg/m3. Answers: 334.1 kg, 3277 N
1–16 At 45° latitude, the gravitational acceleration as a function of elevation z above sea level is given by g ! a $ bz,
where a ! 9.807 m/s2 and b ! 3.32 ' 10$6 s$2. Determine the height above sea level where the weight of an object will decrease by 1 percent. Answer: 29,539 m 1–17E A 150-lbm astronaut took his bathroom scale (a spring scale) and a beam scale (compares masses) to the moon where the local gravity is g ! 5.48 ft/s2. Determine how much he will weigh (a) on the spring scale and (b) on the beam scale. Answers: (a) 25.5 lbf; (b) 150 lbf 1–18 The acceleration of high-speed aircraft is sometimes expressed in g’s (in multiples of the standard acceleration of gravity). Determine the net upward force, in N, that a 90-kg man would experience in an aircraft whose acceleration is 6 g’s. 1–19
A 5-kg rock is thrown upward with a force of 150 N at a location where the local gravitational acceleration is 9.79 m/s2. Determine the acceleration of the rock, in m/s2. 1–20
Solve Prob. 1–19 using EES (or other) software. Print out the entire solution, including the numerical results with proper units. 1–21 The value of the gravitational acceleration g decreases with elevation from 9.807 m/s2 at sea level to 9.767 m/s2 at an altitude of 13,000 m, where large passenger planes cruise. Determine the percent reduction in the weight of an airplane cruising at 13,000 m relative to its weight at sea level.
Modeling and Solving Engineering Problems 1–22C What is the difference between precision and accuracy? Can a measurement be very precise but inaccurate? Explain. 1–23C What is the difference between the analytical and experimental approach to engineering problems? Discuss the advantages and disadvantages of each approach. 1–24C What is the importance of modeling in engineering? How are the mathematical models for engineering processes prepared? 1–25C When modeling an engineering process, how is the right choice made between a simple but crude and a complex but accurate model? Is the complex model necessarily a better choice since it is more accurate? 1–26C How do the differential equations in the study of a physical problem arise?
* Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
1–27C What is the value of the engineering software packages in (a) engineering education and (b) engineering practice? 1–28
Determine a positive real root of this equation using EES: 2x 3 $ 10x 0.5 $ 3x ! $3
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33 CHAPTER 1
1–29
Solve this system of two equations with two unknowns using EES:
x 3 $ y 2 ! 7.75 3xy & y ! 3.5 1–30
Solve this system of three equations with three unknowns using EES:
2x $ y & z ! 5 3x 2 & 2y ! z & 2 xy & 2z ! 8 1–31
Solve this system of three equations with three unknowns using EES:
x 2y $ z ! 1 x $ 3y
0.5
& xz ! $2
x&y$z!2 Review Problems 1–32 The weight of bodies may change somewhat from one location to another as a result of the variation of the gravita-
tional acceleration g with elevation. Accounting for this variation using the relation in Prob. 1–16, determine the weight of an 80-kg person at sea level (z ! 0), in Denver (z ! 1610 m), and on the top of Mount Everest (z ! 8848 m). 1–33 A man goes to a traditional market to buy a steak for dinner. He finds a 12-oz steak (1 lbm = 16 oz) for $3.15. He then goes to the adjacent international market and finds a 320-g steak of identical quality for $2.80. Which steak is the better buy? 1–34 The reactive force developed by a jet engine to push an airplane forward is called thrust, and the thrust developed by the engine of a Boeing 777 is about 85,000 lbf. Express this thrust in N and kgf.
Design and Essay Problem 1–35 Write an essay on the various mass- and volume-measurement devices used throughout history. Also, explain the development of the modern units for mass and volume.
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CHAPTER
2
PROPERTIES OF FLUIDS
I
n this chapter, we discuss properties that are encountered in the analysis of fluid flow. First we discuss intensive and extensive properties and define density and specific gravity. This is followed by a discussion of the properties vapor pressure, energy and its various forms, the specific heats of ideal gases and incompressible substances, and the coefficient of compressibility. Then we discuss the property viscosity, which plays a dominant role in most aspects of fluid flow. Finally, we present the property surface tension and determine the capillary rise from static equilibrium conditions. The property pressure is discussed in Chap. 3 together with fluid statics.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
Have a working knowledge of the basic properties of fluids and understand the continuum approximation Have a working knowledge of viscosity and the consequences of the frictional effects it causes in fluid flow Calculate the capillary rises and drops due to the surface tension effect
35
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36 FLUID MECHANICS
2–1 m V T P ρ
–12 m –12 V T P ρ
–12 m –12 V T P ρ
Extensive properties Intensive properties
FIGURE 2–1 Criteria to differentiate intensive and extensive properties.
■
INTRODUCTION
Any characteristic of a system is called a property. Some familiar properties are pressure P, temperature T, volume V, and mass m. The list can be extended to include less familiar ones such as viscosity, thermal conductivity, modulus of elasticity, thermal expansion coefficient, electric resistivity, and even velocity and elevation. Properties are considered to be either intensive or extensive. Intensive properties are those that are independent of the mass of a system, such as temperature, pressure, and density. Extensive properties are those whose values depend on the size—or extent—of the system. Total mass, total volume V, and total momentum are some examples of extensive properties. An easy way to determine whether a property is intensive or extensive is to divide the system into two equal parts with an imaginary partition, as shown in Fig. 2–1. Each part will have the same value of intensive properties as the original system, but half the value of the extensive properties. Generally, uppercase letters are used to denote extensive properties (with mass m being a major exception), and lowercase letters are used for intensive properties (with pressure P and temperature T being the obvious exceptions). Extensive properties per unit mass are called specific properties. Some examples of specific properties are specific volume (v ! V/m) and specific total energy (e ! E/m). The state of a system is described by its properties. But we know from experience that we do not need to specify all the properties in order to fix a state. Once the values of a sufficient number of properties are specified, the rest of the properties assume certain values. That is, specifying a certain number of properties is sufficient to fix a state. The number of properties required to fix the state of a system is given by the state postulate: The state of a simple compressible system is completely specified by two independent, intensive properties. Two properties are independent if one property can be varied while the other one is held constant. Not all properties are independent, and some are defined in terms of others, as explained in Section 2–2.
Continuum
Matter is made up of atoms that are widely spaced in the gas phase. Yet it is very convenient to disregard the atomic nature of a substance and view it as a continuous, homogeneous matter with no holes, that is, a continuum. The continuum idealization allows us to treat properties as point functions and to assume that the properties vary continually in space with no jump discontinuities. This idealization is valid as long as the size of the system we deal with is large relative to the space between the molecules. This is the case in practically all problems, except some specialized ones. The continuum idealization is implicit in many statements we make, such as “the density of water in a glass is the same at any point.” To have a sense of the distances involved at the molecular level, consider a container filled with oxygen at atmospheric conditions. The diameter of the oxygen molecule is about 3 " 10#10 m and its mass is 5.3 " 10#26 kg. Also, the mean free path of oxygen at 1 atm pressure and 20°C is 6.3 " 10#8 m. That is, an oxygen molecule travels, on average, a distance of 6.3 " 10#8 m (about 200 times its diameter) before it collides with another molecule.
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Also, there are about 2.5 " 1016 molecules of oxygen in the tiny volume of 1 mm3 at 1 atm pressure and 20°C (Fig. 2–2). The continuum model is applicable as long as the characteristic length of the system (such as its diameter) is much larger than the mean free path of the molecules. At very high vacuums or very high elevations, the mean free path may become large (for example, it is about 0.1 m for atmospheric air at an elevation of 100 km). For such cases the rarefied gas flow theory should be used, and the impact of individual molecules should be considered. In this text we limit our consideration to substances that can be modeled as a continuum.
2–2
■
O2
1 atm, 20°C
3 × 1016 molecules/mm3
VOID
DENSITY AND SPECIFIC GRAVITY
Density is defined as mass per unit volume (Fig. 2–3). That is, Density:
m r! V
(kg/m3)
(2–1)
The reciprocal of density is the specific volume v, which is defined as volume per unit mass. That is, v ! V/m ! 1/r. For a differential volume element of mass dm and volume dV, density can be expressed as r ! dm/dV. The density of a substance, in general, depends on temperature and pressure. The density of most gases is proportional to pressure and inversely proportional to temperature. Liquids and solids, on the other hand, are essentially incompressible substances, and the variation of their density with pressure is usually negligible. At 20°C, for example, the density of water changes from 998 kg/m3 at 1 atm to 1003 kg/m3 at 100 atm, a change of just 0.5 percent. The density of liquids and solids depends more strongly on temperature than it does on pressure. At 1 atm, for example, the density of water changes from 998 kg/m3 at 20°C to 975 kg/m3 at 75°C, a change of 2.3 percent, which can still be neglected in many engineering analyses. Sometimes the density of a substance is given relative to the density of a well-known substance. Then it is called specific gravity, or relative density, and is defined as the ratio of the density of a substance to the density of some standard substance at a specified temperature (usually water at 4°C, for which rH2 O ! 1000 kg/m3). That is, Specific gravity:
SG !
r rH2O
(2–2)
Note that the specific gravity of a substance is a dimensionless quantity. However, in SI units, the numerical value of the specific gravity of a substance is exactly equal to its density in g/cm3 or kg/L (or 0.001 times the density in kg/m3) since the density of water at 4°C is 1 g/cm3 ! 1 kg/L ! 1000 kg/m3. The specific gravity of mercury at 0°C, for example, is 13.6. Therefore, its density at 0°C is 13.6 g/cm3 ! 13.6 kg/L ! 13,600 kg/m3. The specific gravities of some substances at 0°C are given in Table 2–1. Note that substances with specific gravities less than 1 are lighter than water, and thus they would float on water. The weight of a unit volume of a substance is called specific weight and is expressed as Specific weight:
gs ! rg
where g is the gravitational acceleration.
(N/m3)
(2–3)
FIGURE 2–2 Despite the large gaps between molecules, a substance can be treated as a continuum because of the very large number of molecules even in an extremely small volume.
V = 12 m 3 m = 3 kg
ρ = 0.25 kg/m 3 1 3 v =– ρ = 4 m /kg
FIGURE 2–3 Density is mass per unit volume; specific volume is volume per unit mass. TA B L E 2 – 1 Specific gravities of some substances at 0°C Substance
SG
Water Blood Seawater Gasoline Ethyl alcohol Mercury Wood Gold Bones Ice Air (at 1 atm)
1.0 1.05 1.025 0.7 0.79 13.6 0.3–0.9 19.2 1.7–2.0 0.92 0.0013
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38 FLUID MECHANICS
Recall from Chap. 1 that the densities of liquids are essentially constant, and thus they can often be approximated as being incompressible substances during most processes without sacrificing much in accuracy.
Density of Ideal Gases
Property tables provide very accurate and precise information about the properties, but sometimes it is convenient to have some simple relations among the properties that are sufficiently general and accurate. Any equation that relates the pressure, temperature, and density (or specific volume) of a substance is called an equation of state. The simplest and best-known equation of state for substances in the gas phase is the ideal-gas equation of state, expressed as Pv ! RT
or
P ! rRT
(2–4)
where P is the absolute pressure, v is the specific volume, T is the thermodynamic (absolute) temperature, r is the density, and R is the gas constant. The gas constant R is different for each gas and is determined from R ! Ru /M, where Ru is the universal gas constant whose value is Ru ! 8.314 kJ/kmol · K ! 1.986 Btu/lbmol · R, and M is the molar mass (also called molecular weight) of the gas. The values of R and M for several substances are given in Table A–1. The thermodynamic temperature scale in the SI is the Kelvin scale, and the temperature unit on this scale is the kelvin, designated by K. In the English system, it is the Rankine scale, and the temperature unit on this scale is the rankine, R. Various temperature scales are related to each other by T(K) ! T($C) % 273.15
(2–5)
T(R) ! T($F) % 459.67
(2–6)
It is common practice to round the constants 273.15 and 459.67 to 273 and 460, respectively. Equation 2–4 is called the ideal-gas equation of state, or simply the ideal-gas relation, and a gas that obeys this relation is called an ideal gas. For an ideal gas of volume V, mass m, and number of moles N ! m/M, the ideal-gas equation of state can also be written as PV ! mRT or PV ! NRuT. For a fixed mass m, writing the ideal-gas relation twice and simplifying, the properties of an ideal gas at two different states are related to each other by P1V1/T1 ! P2V2/T2. An ideal gas is a hypothetical substance that obeys the relation Pv ! RT. It has been experimentally observed that the ideal-gas relation closely approximates the P-v-T behavior of real gases at low densities. At low pressures and high temperatures, the density of a gas decreases and the gas behaves like an ideal gas. In the range of practical interest, many familiar gases such as air, nitrogen, oxygen, hydrogen, helium, argon, neon, and krypton and even heavier gases such as carbon dioxide can be treated as ideal gases with negligible error (often less than 1 percent). Dense gases such as water vapor in steam power plants and refrigerant vapor in refrigerators, however, should not be treated as ideal gases since they usually exist at a state near saturation.
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39 CHAPTER 2
EXAMPLE 2–1
Density, Specific Gravity, and Mass of Air in a Room
Determine the density, specific gravity, and mass of the air in a room whose dimensions are 4 m ! 5 m ! 6 m at 100 kPa and 25°C (Fig. 2–4).
Solution The density, specific gravity, and mass of the air in a room are to be determined. Assumptions At specified conditions, air can be treated as an ideal gas. Properties The gas constant of air is R " 0.287 kPa ! m3/kg ! K. Analysis The density of air is determined from the ideal-gas relation P " rRT to be
r"
P 100 kPa " " 1.17 kg/m3 RT (0.287 kPa # m3/kg # K)(25 $ 273) K
Then the specific gravity of air becomes
SG "
r r H2O
"
1.17 kg/m3 1000 kg/m3
" 0.00117
Finally, the volume and the mass of air in the room are
V " (4 m)(5 m)(6 m) " 120 m3 m " rV " (1.17 kg/m3)(120 m3) " 140 kg Discussion Note that we converted the temperature to the unit K from °C before using it in the ideal-gas relation.
2–3
■
VAPOR PRESSURE AND CAVITATION
It is well-established that temperature and pressure are dependent properties for pure substances during phase-change processes, and there is one-to-one correspondence between temperatures and pressures. At a given pressure, the temperature at which a pure substance changes phase is called the saturation temperature Tsat. Likewise, at a given temperature, the pressure at which a pure substance changes phase is called the saturation pressure Psat. At an absolute pressure of 1 standard atmosphere (1 atm or 101.325 kPa), for example, the saturation temperature of water is 100°C. Conversely, at a temperature of 100°C, the saturation pressure of water is 1 atm. The vapor pressure Pv of a pure substance is defined as the pressure exerted by its vapor in phase equilibrium with its liquid at a given temperature. Pv is a property of the pure substance, and turns out to be identical to the saturation pressure Psat of the liquid (Pv " Psat). We must be careful not to confuse vapor pressure with partial pressure. Partial pressure is defined as the pressure of a gas or vapor in a mixture with other gases. For example, atmospheric air is a mixture of dry air and water vapor, and atmospheric pressure is the sum of the partial pressure of dry air and the partial pressure of water vapor. The partial pressure of water vapor constitutes a small fraction (usually under 3 percent) of the atmospheric pressure since air is mostly nitrogen and oxygen. The partial pressure of a vapor must be less than or equal to the vapor pressure if there is no liquid present. However, when both vapor and liquid are present and the system is in phase equilibrium, the partial pressure of the vapor must equal the vapor pressure, and the system is said to be saturated. The rate of evaporation from open water bodies such as
4m
6m
AIR P = 100 kPa T = 25°C
5m
FIGURE 2–4 Schematic for Example 2–1.
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40 FLUID MECHANICS
TABLE 2–2 Saturation (or vapor) pressure of water at various temperatures Temperature T, °C &10 &5 0 5 10 15 20 25 30 40 50 100 150 200 250 300
Saturation Pressure Psat, kPa 0.260 0.403 0.611 0.872 1.23 1.71 2.34 3.17 4.25 7.38 12.35 101.3 (1 atm) 475.8 1554 3973 8581
FIGURE 2–5 Cavitation damage on a 16-mm by 23-mm aluminum sample tested at 60 m/s for 2.5 h. The sample was located at the cavity collapse region downstream of a cavity generator specifically designed to produce high damage potential. Photograph by David Stinebring, ARL/Pennsylvania State University. Used by permission.
lakes is controlled by the difference between the vapor pressure and the partial pressure. For example, the vapor pressure of water at 20°C is 2.34 kPa. Therefore, a bucket of water at 20°C left in a room with dry air at 1 atm will continue evaporating until one of two things happens: the water evaporates away (there is not enough water to establish phase equilibrium in the room), or the evaporation stops when the partial pressure of the water vapor in the room rises to 2.34 kPa at which point phase equilibrium is established. For phase-change processes between the liquid and vapor phases of a pure substance, the saturation pressure and the vapor pressure are equivalent since the vapor is pure. Note that the pressure value would be the same whether it is measured in the vapor or liquid phase (provided that it is measured at a location close to the liquid–vapor interface to avoid the hydrostatic effects). Vapor pressure increases with temperature. Thus, a substance at higher temperatures boils at higher pressures. For example, water boils at 134°C in a pressure cooker operating at 3 atm absolute pressure, but it boils at 93°C in an ordinary pan at a 2000-m elevation, where the atmospheric pressure is 0.8 atm. The saturation (or vapor) pressures are given in Appendices 1 and 2 for various substances. A mini table for water is given in Table 2–2 for easy reference. The reason for our interest in vapor pressure is the possibility of the liquid pressure in liquid-flow systems dropping below the vapor pressure at some locations, and the resulting unplanned vaporization. For example, water at 10°C will flash into vapor and form bubbles at locations (such as the tip regions of impellers or suction sides of pumps) where the pressure drops below 1.23 kPa. The vapor bubbles (called cavitation bubbles since they form “cavities” in the liquid) collapse as they are swept away from the lowpressure regions, generating highly destructive, extremely high-pressure waves. This phenomenon, which is a common cause for drop in performance and even the erosion of impeller blades, is called cavitation, and it is an important consideration in the design of hydraulic turbines and pumps (Fig. 2–5). Cavitation must be avoided (or at least minimized) in flow systems since it reduces performance, generates annoying vibrations and noise, and causes damage to equipment. The pressure spikes resulting from the large number of bubbles collapsing near a solid surface over a long period of time may cause erosion, surface pitting, fatigue failure, and the eventual destruction of the components or machinery. The presence of cavitation in a flow system can be sensed by its characteristic tumbling sound. EXAMPLE 2–2
Minimum Pressure to Avoid Cavitation
In a water distribution system, the temperature of water is observed to be as high as 30°C. Determine the minimum pressure allowed in the system to avoid cavitation.
SOLUTION The minimum pressure in a water distribution system to avoid cavitation is to be determined. Properties The vapor pressure of water at 30°C is 4.25 kPa. Analysis To avoid cavitation, the pressure anywhere in the flow should not be allowed to drop below the vapor (or saturation) pressure at the given temperature. That is, Pmin " Psat@30%C " 4.25 kPa
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41 CHAPTER 2
Therefore, the pressure should be maintained above 4.25 kPa everywhere in the flow. Discussion Note that the vapor pressure increases with increasing temperature, and thus the risk of cavitation is greater at higher fluid temperatures.
2–4
■
ENERGY AND SPECIFIC HEATS
Energy can exist in numerous forms such as thermal, mechanical, kinetic, potential, electrical, magnetic, chemical, and nuclear, and their sum constitutes the total energy E (or e on a unit mass basis) of a system. The forms of energy related to the molecular structure of a system and the degree of the molecular activity are referred to as the microscopic energy. The sum of all microscopic forms of energy is called the internal energy of a system, and is denoted by U (or u on a unit mass basis). The macroscopic energy of a system is related to motion and the influence of some external effects such as gravity, magnetism, electricity, and surface tension. The energy that a system possesses as a result of its motion relative to some reference frame is called kinetic energy. When all parts of a system move with the same velocity, the kinetic energy per unit mass is expressed as ke ! V 2/2 where V denotes the velocity of the system relative to some fixed reference frame. The energy that a system possesses as a result of its elevation in a gravitational field is called potential energy and is expressed on a per-unit mass basis as pe ! gz where g is the gravitational acceleration and z is the elevation of the center of gravity of a system relative to some arbitrarily selected reference plane. In daily life, we frequently refer to the sensible and latent forms of internal energy as heat, and we talk about the heat content of bodies. In engineering, however, those forms of energy are usually referred to as thermal energy to prevent any confusion with heat transfer. The international unit of energy is the joule (J) or kilojoule (1 kJ ! 1000 J). In the English system, the unit of energy is the British thermal unit (Btu), which is defined as the energy needed to raise the temperature of 1 lbm of water at 68°F by 1°F. The magnitudes of kJ and Btu are almost identical (1 Btu ! 1.0551 kJ). Another well-known unit of energy is the calorie (1 cal ! 4.1868 J), which is defined as the energy needed to raise the temperature of 1 g of water at 14.5°C by 1°C. In the analysis of systems that involve fluid flow, we frequently encounter the combination of properties u and Pv. For convenience, this combination is called enthalpy h. That is, Enthalpy:
h ! u % Pv ! u %
P r
Flowing fluid
Energy = h
(2–7)
where P/r is the flow energy, also called the flow work, which is the energy per unit mass needed to move the fluid and maintain flow. In the energy analysis of flowing fluids, it is convenient to treat the flow energy as part of the energy of the fluid and to represent the microscopic energy of a fluid stream by enthalpy h (Fig. 2–6). Note that enthalpy is a quantity per unit mass, and thus it is a specific property. In the absence of such effects as magnetic, electric, and surface tension, a system is called a simple compressible system. The total energy of a simple
Stationary fluid
Energy = u
FIGURE 2–6 The internal energy u represents the microscopic energy of a nonflowing fluid per unit mass, whereas enthalpy h represents the microscopic energy of a flowing fluid per unit mass.
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42 FLUID MECHANICS
compressible system consists of three parts: internal, kinetic, and potential energies. On a unit-mass basis, it is expressed as e ! u % ke % pe. The fluid entering or leaving a control volume possesses an additional form of energy—the flow energy P/r. Then the total energy of a flowing fluid on a unit-mass basis becomes eflowing ! P/r % e ! h % ke % pe ! h %
P1
V2 % gz 2
(kJ/kg)
(2–8)
where h ! P/r % u is the enthalpy, V is the velocity, and z is the elevation of the system relative to some external reference point. By using the enthalpy instead of the internal energy to represent the energy of a flowing fluid, one does not need to be concerned about the flow work. The energy associated with pushing the fluid is automatically taken care of by enthalpy. In fact, this is the main reason for defining the property enthalpy. The differential and finite changes in the internal energy and enthalpy of an ideal gas can be expressed in terms of the specific heats as du ! cv dT
and
dh ! cp dT
(2–9)
where cv and cp are the constant-volume and constant-pressure specific heats of the ideal gas. Using specific heat values at the average temperature, the finite changes in internal energy and enthalpy can be expressed approximately as &u ! cv,ave &T
P2 > P1
and
&h ! cp,ave &T
(2–10)
For incompressible substances, the constant-volume and constant-pressure specific heats are identical. Therefore, cp ! cv ! c for liquids, and the change in the internal energy of liquids can be expressed as &u ! cave &T. Noting that r ! constant for incompressible substances, the differentiation of enthalpy h ! u % P/r gives dh ! du % dP/r. Integrating, the enthalpy change becomes &h ! &u % &P/r ! cave &T % &P/r
(2–11)
Therefore, &h ! &u ! cave &T for constant-pressure processes, and &h ! &P/r for constant-temperature processes of liquids.
2–5 FIGURE 2–7 Fluids, like solids, compress when the applied pressure is increased from P1 to P2.
■
COEFFICIENT OF COMPRESSIBILITY
We know from experience that the volume (or density) of a fluid changes with a change in its temperature or pressure. Fluids usually expand as they are heated or depressurized and contract as they are cooled or pressurized. But the amount of volume change is different for different fluids, and we need to define properties that relate volume changes to the changes in pressure and temperature. Two such properties are the bulk modulus of elasticity k and the coefficient of volume expansion b. It is a common observation that a fluid contracts when more pressure is applied on it and expands when the pressure acting on it is reduced (Fig. 2–7). That is, fluids act like elastic solids with respect to pressure. Therefore, in an analogous manner to Young’s modulus of elasticity for solids, it is appropriate to define a coefficient of compressibility k (also called the bulk modulus of compressibility or bulk modulus of elasticity) for fluids as k ! #v a
'P 'P b ! ra b 'v T 'r T
(Pa)
(2–12)
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43 CHAPTER 2
It can also be expressed approximately in terms of finite changes as k!#
&P &P ! &v/v &r/r
(T ! constant)
(2–13)
Noting that &v/v or &r/r is dimensionless, k must have the dimension of pressure (Pa or psi). Also, the coefficient of compressibility represents the change in pressure corresponding to a fractional change in volume or density of the fluid while the temperature remains constant. Then it follows that the coefficient of compressibility of a truly incompressible substance (v ! constant) is infinity. A large value of k indicates that a large change in pressure is needed to cause a small fractional change in volume, and thus a fluid with a large k is essentially incompressible. This is typical for liquids, and explains why liquids are usually considered to be incompressible. For example, the pressure of water at normal atmospheric conditions must be raised to 210 atm to compress it 1 percent, corresponding to a coefficient of compressibility value of k ! 21,000 atm. Small density changes in liquids can still cause interesting phenomena in piping systems such as the water hammer—characterized by a sound that resembles the sound produced when a pipe is “hammered.” This occurs when a liquid in a piping network encounters an abrupt flow restriction (such as a closing valve) and is locally compressed. The acoustic waves produced strike the pipe surfaces, bends, and valves as they propagate and reflect along the pipe, causing the pipe to vibrate and produce the familiar sound. Note that volume and pressure are inversely proportional (volume decreases as pressure is increased and thus ∂P/∂v is a negative quantity), and the negative sign in the definition (Eq. 2–12) ensures that k is a positive quantity. Also, differentiating r ! 1/v gives dr ! #dv/v 2, which can be rearranged as dr dv !# r v
(2–14)
That is, the fractional changes in the specific volume and the density of a fluid are equal in magnitude but opposite in sign. For an ideal gas, P ! rRT and (∂P/∂r)T ! RT ! P/r, and thus kideal gas ! P
(Pa)
(2–15)
Therefore, the coefficient of compressibility of an ideal gas is equal to its absolute pressure, and the coefficient of compressibility of the gas increases with increasing pressure. Substituting k ! P into the definition of the coefficient of compressibility and rearranging gives Ideal gas:
&r &P ! r P
(T ! constant)
(2–16)
Therefore, the percent increase of density of an ideal gas during isothermal compression is equal to the percent increase in pressure. For air at 1 atm pressure, k ! P ! 1 atm and a decrease of 1 percent in volume (&V/V ! #0.01) corresponds to an increase of &P ! 0.01 atm in pressure. But for air at 1000 atm, k ! 1000 atm and a decrease of 1 percent in volume corresponds to an increase of &P ! 10 atm in pressure. Therefore,
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44 FLUID MECHANICS
a small fractional change in the volume of a gas can cause a large change in pressure at very high pressures. The inverse of the coefficient of compressibility is called the isothermal compressibility a and is expressed as a!
1 1 'v 1 'r !# a b ! a b k v 'P T r 'P T
(1/Pa)
(2–17)
The isothermal compressibility of a fluid represents the fractional change in volume or density corresponding to a unit change in pressure.
Coefficient of Volume Expansion
FIGURE 2–8 Natural convection over a woman’s hand. G. S. Settles, Gas Dynamics Lab, Penn State University. Used by permission.
The density of a fluid, in general, depends more strongly on temperature than it does on pressure, and the variation of density with temperature is responsible for numerous natural phenomena such as winds, currents in oceans, rise of plumes in chimneys, the operation of hot-air balloons, heat transfer by natural convection, and even the rise of hot air and thus the phrase “heat rises” (Fig. 2–8). To quantify these effects, we need a property that represents the variation of the density of a fluid with temperature at constant pressure. The property that provides that information is the coefficient of volume expansion (or volume expansivity) b, defined as (Fig. 2–9) b !
∂v
Q ––– ∂T R
21°C 100 kPa 1 kg
(a) A substance with a large b ∂v Q––– ∂T RP 20°C 100 kPa 1 kg
21°C 100 kPa 1 kg
(b) A substance with a small b
FIGURE 2–9 The coefficient of volume expansion is a measure of the change in volume of a substance with temperature at constant pressure.
(1/K)
(2–18)
It can also be expressed approximately in terms of finite changes as
P
20°C 100 kPa 1 kg
1 'r 1 'v a b !# a b r 'T P v 'T P
b"
&r/r &v/v !# &T &T
(at constant P)
(2–19)
A large value of b for a fluid means a large change in density with temperature, and the product b &T represents the fraction of volume change of a fluid that corresponds to a temperature change of &T at constant pressure. It can be shown easily that the volume expansion coefficient of an ideal gas (P ! rRT ) at a temperature T is equivalent to the inverse of the temperature: b ideal gas !
1 T
(1/K)
(2–20)
where T is the absolute temperature. In the study of natural convection currents, the condition of the main fluid body that surrounds the finite hot or cold regions is indicated by the subscript “infinity” to serve as a reminder that this is the value at a distance where the presence of the hot or cold region is not felt. In such cases, the volume expansion coefficient can be expressed approximately as b"#
(r ( # r)/r T( # T
or
r ( # r ! rb(T # T()
(2–21)
where r( is the density and T( is the temperature of the quiescent fluid away from the confined hot or cold fluid pocket.
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45 CHAPTER 2
We will see in Chap. 3 that natural convection currents are initiated by the buoyancy force, which is proportional to the density difference, which is proportional to the temperature difference at constant pressure. Therefore, the larger the temperature difference between the hot or cold fluid pocket and the surrounding main fluid body, the larger the buoyancy force and thus the stronger the natural convection currents. The combined effects of pressure and temperature changes on the volume change of a fluid can be determined by taking the specific volume to be a function of T and P. Differentiating v ! v(T, P) and using the definitions of the compression and expansion coefficients a and b give 'v 'v dv ! a b dT % a b dP ! (b dT # a dP)v 'T P 'P T
(2–22)
Then the fractional change in volume (or density) due to changes in pressure and temperature can be expressed approximately as &r &v ! b &T # a &P !# r v
EXAMPLE 2–3
(2–23)
Variation of Density with Temperature and Pressure
Consider water initially at 20°C and 1 atm. Determine the final density of water (a) if it is heated to 50°C at a constant pressure of 1 atm, and (b) if it is compressed to 100-atm pressure at a constant temperature of 20°C. Take the isothermal compressibility of water to be a ! 4.80 " 10#5 atm#1.
SOLUTION Water at a given temperature and pressure is considered. The densities of water after it is heated and after it is compressed are to be determined. Assumptions 1 The coefficient of volume expansion and the isothermal compressibility of water are constant in the given temperature range. 2 An approximate analysis is performed by replacing differential changes in quantities by finite changes. Properties The density of water at 20°C and 1 atm pressure is r1 ! 998.0 kg/m3. The coefficient of volume expansion at the average temperature of (20 % 50)/2 ! 35°C is b ! 0.337 " 10#3 K#1. The isothermal compressibility of water is given to be a ! 4.80 " 10#5 atm#1. Analysis When differential quantities are replaced by differences and the properties a and b are assumed to be constant, the change in density in terms of the changes in pressure and temperature is expressed approximately as (Eq. 2–23) &r ! ar &P # br &T (a) The change in density due to the change of temperature from 20°C to 50°C at constant pressure is
&r ! #br &T ! #(0.337 " 10 #3 K #1)(998 kg/m3)(50 # 20) K ! #10.0 kg/m3 Noting that &r ! r2 # r1, the density of water at 50°C and 1 atm is
r 2 ! r 1 % &r ! 998.0 % (#10.0) ! 988.0 kg /m3
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46 FLUID MECHANICS 0.00050
which is almost identical to the listed value of 988.1 kg/m3 at 50°C in Table A–3. This is mostly due to b varying with temperature almost linearly, as shown in Fig. 2–10.
b, 1/K
0.00045 0.00040
0.00030
(b) The change in density due to a change of pressure from 1 atm to 100 atm at constant temperature is
0.00025
&r ! ar &P ! (4.80 " 10 #5 atm #1)(998 kg/m3)(100 # 1) atm ! 4.7 kg/m3
0.00035
0.00020 20
25
30
35 40 T, °C
45
50
FIGURE 2–10 The variation of the coefficient of volume expansion of water b with temperature in the range of 20°C to 50°C.
Then the density of water at 100 atm and 20°C becomes
r 2 ! r 1 % &r ! 998.0 % 4.7 ! 1002.7 kg/m3 Discussion Note that the density of water decreases while being heated and increases while being compressed, as expected. This problem can be solved more accurately using differential analysis when functional forms of properties are available.
Data were generated and plotted using EES.
V
2–6
Air
Drag force
V
Water
Drag force
FIGURE 2–11 A fluid moving relative to a body exerts a drag force on the body, partly because of friction caused by viscosity.
■
VISCOSITY
When two solid bodies in contact move relative to each other, a friction force develops at the contact surface in the direction opposite to motion. To move a table on the floor, for example, we have to apply a force to the table in the horizontal direction large enough to overcome the friction force. The magnitude of the force needed to move the table depends on the friction coefficient between the table and the floor. The situation is similar when a fluid moves relative to a solid or when two fluids move relative to each other. We move with relative ease in air, but not so in water. Moving in oil would be even more difficult, as can be observed by the slower downward motion of a glass ball dropped in a tube filled with oil. It appears that there is a property that represents the internal resistance of a fluid to motion or the “fluidity,” and that property is the viscosity. The force a flowing fluid exerts on a body in the flow direction is called the drag force, and the magnitude of this force depends, in part, on viscosity (Fig. 2–11). To obtain a relation for viscosity, consider a fluid layer between two very large parallel plates (or equivalently, two parallel plates immersed in a large body of a fluid) separated by a distance ! (Fig. 2–12). Now a constant parallel force F is applied to the upper plate while the lower plate is held fixed. After the initial transients, it is observed that the upper plate moves continuously under the influence of this force at a constant velocity V. The fluid in contact with the upper plate sticks to the plate surface and moves with it at the same velocity, and the shear stress t acting on this fluid layer is t!
F A
(2–24)
where A is the contact area between the plate and the fluid. Note that the fluid layer deforms continuously under the influence of shear stress. The fluid in contact with the lower plate assumes the velocity of that plate, which is zero (again because of the no-slip condition). In steady laminar
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47 CHAPTER 2
flow, the fluid velocity between the plates varies linearly between 0 and V, and thus the velocity profile and the velocity gradient are y u(y) ! V /
du V ! dy /
and
(2–25)
da N
where y is the vertical distance from the lower plate. During a differential time interval dt, the sides of fluid particles along a vertical line MN rotate through a differential angle db while the upper plate moves a differential distance da ! V dt. The angular displacement or deformation (or shear strain) can be expressed as db " tan b !
da V dt du ! ! dt / / dy
x
Velocity V ! u= 0
M
Velocity profile y u(y) = V !
FIGURE 2–12 The behavior of a fluid in laminar flow between two parallel plates when the upper plate moves with a constant velocity.
(2–26)
Rearranging, the rate of deformation under the influence of shear stress t becomes db du ! dt dy
Force F
N′ u = V
db
y
Area A
(2–27)
Thus we conclude that the rate of deformation of a fluid element is equivalent to the velocity gradient du/dy. Further, it can be verified experimentally that for most fluids the rate of deformation (and thus the velocity gradient) is directly proportional to the shear stress t, t )
db dt
or
t )
du dy
(2–28)
Fluids for which the rate of deformation is proportional to the shear stress are called Newtonian fluids after Sir Isaac Newton, who expressed it first in 1687. Most common fluids such as water, air, gasoline, and oils are Newtonian fluids. Blood and liquid plastics are examples of non-Newtonian fluids. In one-dimensional shear flow of Newtonian fluids, shear stress can be expressed by the linear relationship Oil 2
(N/m )
m =
where the constant of proportionality m is called the coefficient of viscosity or the dynamic (or absolute) viscosity of the fluid, whose unit is kg/m · s, or equivalently, N · s/m2 (or Pa ! s where Pa is the pressure unit pascal). A common viscosity unit is poise, which is equivalent to 0.1 Pa ! s (or centipoise, which is one-hundredth of a poise). The viscosity of water at 20°C is 1 centipoise, and thus the unit centipoise serves as a useful reference. A plot of shear stress versus the rate of deformation (velocity gradient) for a Newtonian fluid is a straight line whose slope is the viscosity of the fluid, as shown in Fig. 2–13. Note that viscosity is independent of the rate of deformation. The shear force acting on a Newtonian fluid layer (or, by Newton’s third law, the force acting on the plate) is Shear force:
F ! tA ! mA
Viscosity = Slope
(2–29)
du dy
(N)
(2–30)
Shear stress, t
Shear stress:
du t ! m dy
a
t du / dy
=
a b
Water
b
Air Rate of deformation, du/dy
FIGURE 2–13 The rate of deformation (velocity gradient) of a Newtonian fluid is proportional to shear stress, and the constant of proportionality is the viscosity.
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48 FLUID MECHANICS Bingham plastic Shear stress, t
Pseudoplastic Newtonian
Dilatant
Rate of deformation, du/dy
FIGURE 2–14 Variation of shear stress with the rate of deformation for Newtonian and non-Newtonian fluids (the slope of a curve at a point is the apparent viscosity of the fluid at that point).
Air at 20°C and 1 atm: m = 1.83 × 10–5 kg/m ⋅ s * = 1.52 × 10–5 m2/s Air at 20°C and 4 atm: m = 1.83 × 10–5 kg/m ⋅ s * = 0.380 × 10–5 m2/s
FIGURE 2–15 Dynamic viscosity, in general, does not depend on pressure, but kinematic viscosity does.
where again A is the contact area between the plate and the fluid. Then the force F required to move the upper plate in Fig. 2–12 at a constant velocity of V while the lower plate remains stationary is F ! mA
V /
(N)
(2–31)
This relation can alternately be used to calculate m when the force F is measured. Therefore, the experimental setup just described can be used to measure the viscosity of fluids. Note that under identical conditions, the force F will be very different for different fluids. For non-Newtonian fluids, the relationship between shear stress and rate of deformation is not linear, as shown in Fig. 2–14. The slope of the curve on the t versus du/dy chart is referred to as the apparent viscosity of the fluid. Fluids for which the apparent viscosity increases with the rate of deformation (such as solutions with suspended starch or sand) are referred to as dilatant or shear thickening fluids, and those that exhibit the opposite behavior (the fluid becoming less viscous as it is sheared harder, such as some paints, polymer solutions, and fluids with suspended particles) are referred to as pseudoplastic or shear thinning fluids. Some materials such as toothpaste can resist a finite shear stress and thus behave as a solid, but deform continuously when the shear stress exceeds the yield stress and thus behave as a fluid. Such materials are referred to as Bingham plastics after E. C. Bingham, who did pioneering work on fluid viscosity for the U.S. National Bureau of Standards in the early twentieth century. In fluid mechanics and heat transfer, the ratio of dynamic viscosity to density appears frequently. For convenience, this ratio is given the name kinematic viscosity n and is expressed as n ! m/r. Two common units of kinematic viscosity are m2/s and stoke (1 stoke ! 1 cm2/s ! 0.0001 m2/s). In general, the viscosity of a fluid depends on both temperature and pressure, although the dependence on pressure is rather weak. For liquids, both the dynamic and kinematic viscosities are practically independent of pressure, and any small variation with pressure is usually disregarded, except at extremely high pressures. For gases, this is also the case for dynamic viscosity (at low to moderate pressures), but not for kinematic viscosity since the density of a gas is proportional to its pressure (Fig. 2–15). The viscosity of a fluid is a measure of its “resistance to deformation.” Viscosity is due to the internal frictional force that develops between different layers of fluids as they are forced to move relative to each other. Viscosity is caused by the cohesive forces between the molecules in liquids and by the molecular collisions in gases, and it varies greatly with temperature. The viscosity of liquids decreases with temperature, whereas the viscosity of gases increases with temperature (Fig. 2–16). This is because in a liquid the molecules possess more energy at higher temperatures, and they can oppose the large cohesive intermolecular forces more strongly. As a result, the energized liquid molecules can move more freely. In a gas, on the other hand, the intermolecular forces are negligible, and the gas molecules at high temperatures move randomly at higher velocities. This results in more molecular collisions per unit volume per unit time and therefore in greater resistance to flow. The viscosity of a fluid is directly
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49 CHAPTER 2
related to the pumping power needed to transport a fluid in a pipe or to move a body (such as a car in air or a submarine in the sea) through a fluid. The kinetic theory of gases predicts the viscosity of gases to be proportional to the square root of temperature. That is, m gas ) 1T . This prediction is confirmed by practical observations, but deviations for different gases need to be accounted for by incorporating some correction factors. The viscosity of gases is expressed as a function of temperature by the Sutherland correlation (from The U.S. Standard Atmosphere) as
Viscosity
Liquids
1/2
Gases:
m!
aT 1 % b/T
(2–32)
where T is absolute temperature and a and b are experimentally determined constants. Note that measuring viscosities at two different temperatures is sufficient to determine these constants. For air, the values of these constants are a ! 1.458 " 10#6 kg/(m ! s ! K1/2) and b ! 110.4 K at atmospheric conditions. The viscosity of gases is independent of pressure at low to moderate pressures (from a few percent of 1 atm to several atm). But viscosity increases at high pressures due to the increase in density. For liquids, the viscosity is approximated as Liquids:
b/(T#c)
m ! a10
(2–33)
where again T is absolute temperature and a, b, and c are experimentally determined constants. For water, using the values a ! 2.414 " 10#5 N ! s/m2, b ! 247.8 K, and c ! 140 K results in less than 2.5 percent error in viscosity in the temperature range of 0°C to 370°C (Touloukian et al., 1975). Consider a fluid layer of thickness ! within a small gap between two concentric cylinders, such as the thin layer of oil in a journal bearing. The gap between the cylinders can be modeled as two parallel flat plates separated by a fluid. Noting that torque is T ! FR (force times the moment arm, which is the radius R of the inner cylinder in this case), the tangential velocity is V ! vR (angular velocity times the radius), and taking the wetted surface area of the inner cylinder to be A ! 2pRL by disregarding the shear stress acting on the two ends of the inner cylinder, torque can be expressed as T ! FR ! m
# 2pR3vL 4p 2R3nL !m / /
(2–34)
. where L is the length of the cylinder and n is the number of revolutions per unit time, which is usually expressed in rpm (revolutions per minute). Note that the angular distance traveled during one rotation is 2p rad, and thus the . relation between the angular velocity in rad/min and the rpm is v ! 2pn. Equation 2–34 can be used to calculate the viscosity of a fluid by measuring torque at a specified angular velocity. Therefore, two concentric cylinders can be used as a viscometer, a device that measures viscosity. The viscosities of some fluids at room temperature are listed in Table 2–3. They are plotted against temperature in Fig. 2–17. Note that the viscosities of different fluids differ by several orders of magnitude. Also note that it is more difficult to move an object in a higher-viscosity fluid such as engine oil than it is in a lower-viscosity fluid such as water. Liquids, in general, are much more viscous than gases.
Gases
Temperature
FIGURE 2–16 The viscosity of liquids decreases and the viscosity of gases increases with temperature. TA B L E 2 – 3 Dynamic viscosities of some fluids at 1 atm and 20°C (unless otherwise stated) Fluid Glycerin: #20°C 0°C 20°C 40°C Engine oil: SAE 10W SAE 10W30 SAE 30 SAE 50 Mercury Ethyl alcohol Water: 0°C 20°C 100°C (liquid) 100°C (vapor) Blood, 37$C Gasoline Ammonia Air Hydrogen, 0°C
Dynamic Viscosity m, kg/m ! s 134.0 10.5 1.52 0.31 0.10 0.17 0.29 0.86 0.0015 0.0012 0.0018 0.0010 0.00028 0.000012 0.00040 0.00029 0.00015 0.000018 0.0000088
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50 FLUID MECHANICS 0.5 0.4 0.3 0.2 0.1
Glycerin
0.06 0.04 0.03 0.02 Absolute viscosity m, N ⋅ s/m2
Castor oil
SAE 10 oil
SAE 30 oil Crude oil (SG 0.86)
0.01 6 4 3 2
Kerosene Aniline Mercury
1 × 10 –3
Carbon tetrachloride
6 4 3 2
Benzene
Ethyl alcohol Water Gasoline (SG 0.68)
1 × 10 –4
FIGURE 2–17 The variation of dynamic (absolute) viscosities of common fluids with temperature at 1 atm (1 N ! s/m2 ! 1 kg/m ! s ! 0.020886 lbf ! s/ft2). F. M. White, Fluid Mechanics 4e. Copyright © 1999 The McGraw-Hill Companies, Inc. Used by permission.
6 4 3 2
Helium Air
1 × 10 –5
Hydrogen
5 –20
EXAMPLE 2–4
Stationary cylinder
R
!
n⋅ = 300 rpm Shaft Fluid
FIGURE 2–18 Schematic for Example 2–4.
Carbon Dioxide
0
20
40 60 80 Temperature, °C
100
120
Determining the Viscosity of a Fluid
The viscosity of a fluid is to be measured by a viscometer constructed of two 40-cm-long concentric cylinders (Fig. 2–18). The outer diameter of the inner cylinder is 12 cm, and the gap between the two cylinders is 0.15 cm. The inner cylinder is rotated at 300 rpm, and the torque is measured to be 1.8 N ! m. Determine the viscosity of the fluid.
SOLUTION The torque and the rpm of a double cylinder viscometer are given. The viscosity of the fluid is to be determined. Assumptions 1 The inner cylinder is completely submerged in oil. 2 The viscous effects on the two ends of the inner cylinder are negligible. Analysis The velocity profile is linear only when the curvature effects are negligible, and the profile can be approximated as being linear in this case since !/R ++ 1. Solving Eq. 2–34 for viscosity and substituting the given values, the viscosity of the fluid is determined to be m!
(1.8 N , m)(0.0015 m) T/ ! ! 0.158 N , s /m2 2 3# 2 4p R nL 4p (0.06 m)3(300/60 1/s)(0.4 m)
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51 CHAPTER 2
Discussion Viscosity is a strong function of temperature, and a viscosity value without a corresponding temperature is of little value. Therefore, the temperature of the fluid should have also been measured during this experiment, and reported with this calculation.
2–7
■
SURFACE TENSION AND CAPILLARY EFFECT
It is often observed that a drop of blood forms a hump on a horizontal glass; a drop of mercury forms a near-perfect sphere and can be rolled just like a steel ball over a smooth surface; water droplets from rain or dew hang from branches or leaves of trees; a liquid fuel injected into an engine forms a mist of spherical droplets; water dripping from a leaky faucet falls as spherical droplets; a soap bubble released into the air forms a spherical shape; and water beads up into small drops on flower petals (Fig. 2–19). In these and other observances, liquid droplets behave like small spherical balloons filled with the liquid, and the surface of the liquid acts like a stretched elastic membrane under tension. The pulling force that causes this tension acts parallel to the surface and is due to the attractive forces between the molecules of the liquid. The magnitude of this force per unit length is called surface tension ss and is usually expressed in the unit N/m (or lbf/ft in English units). This effect is also called surface energy and is expressed in the equivalent unit of N ! m/m2 or J/m2. In this case, ss represents the stretching work that needs to be done to increase the surface area of the liquid by a unit amount. To visualize how surface tension arises, we present a microscopic view in Fig. 2–20 by considering two liquid molecules, one at the surface and one deep within the liquid body. The attractive forces applied on the interior molecule by the surrounding molecules balance each other because of symmetry. But the attractive forces acting on the surface molecule are not symmetric, and the attractive forces applied by the gas molecules above are usually very small. Therefore, there is a net attractive force acting on the molecule at the surface of the liquid, which tends to pull the molecules on the surface toward the interior of the liquid. This force is balanced by the repulsive forces from the molecules below the surface that are being compressed. The resulting compression effect causes the liquid to minimize its surface area. This is the reason for the tendency of the liquid droplets to attain a spherical shape, which has the minimum surface area for a given volume. You also may have observed, with amusement, that some insects can land on water or even walk on water (Fig. 2–19b) and that small steel needles can float on water. These phenomena are again made possible by surface tension that balances the weights of these objects. To understand the surface tension effect better, consider a liquid film (such as the film of a soap bubble) suspended on a U-shaped wire frame with a movable side (Fig. 2–21). Normally, the liquid film tends to pull the movable wire inward in order to minimize its surface area. A force F needs to be applied on the movable wire in the opposite direction to balance this pulling effect. The thin film in the device has two surfaces (the top and bottom
(a)
(b)
FIGURE 2–19 Some consequences of surface tension. (a) © Pegasus/Visuals Unlimited. (b) © Dennis Drenner/Visuals Unlimited.
A molecule on the surface
A molecule inside the liquid
FIGURE 2–20 Attractive forces acting on a liquid molecule at the surface and deep inside the liquid.
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52 FLUID MECHANICS Rigid wire frame Surface of film
Movable wire F
b ∆x x
σs Liquid film
σs
F Wire
FIGURE 2–21 Stretching a liquid film with a U-shaped wire, and the forces acting on the movable wire of length b.
TA B L E 2 – 4 Surface tension of some fluids in air at 1 atm and 20°C (unless otherwise stated) Fluid Water: 0°C 20°C 100°C 300°C Glycerin SAE 30 oil Mercury Ethyl alcohol Blood, 37°C Gasoline Ammonia Soap solution Kerosene
Surface Tension ss, N/m* 0.076 0.073 0.059 0.014 0.063 0.035 0.440 0.023 0.058 0.022 0.021 0.025 0.028
* Multiply by 0.06852 to convert to lbf/ft.
surfaces) exposed to air, and thus the length along which the tension acts in this case is 2b. Then a force balance on the movable wire gives F ! 2bss, and thus the surface tension can be expressed as ss !
F 2b
(2–35)
Note that for b ! 0.5 m, the force F measured (in N) is simply the surface tension in N/m. An apparatus of this kind with sufficient precision can be used to measure the surface tension of various fluids. In the U-shaped wire, the force F remains constant as the movable wire is pulled to stretch the film and increase its surface area. When the movable wire is pulled a distance &x, the surface area increases by &A ! 2b &x, and the work done W during this stretching process is W ! Force " Distance ! F &x ! 2bss &x ! ss &A
since the force remains constant in this case. This result can also be interpreted as the surface energy of the film is increased by an amount ss &A during this stretching process, which is consistent with the alternative interpretation of ss as surface energy. This is similar to a rubber band having more potential (elastic) energy after it is stretched further. In the case of liquid film, the work is used to move liquid molecules from the interior parts to the surface against the attraction forces of other molecules. Therefore, surface tension also can be defined as the work done per unit increase in the surface area of the liquid. The surface tension varies greatly from substance to substance, and with temperature for a given substance, as shown in Table 2–4. At 20°C, for example, the surface tension is 0.073 N/m for water and 0.440 N/m for mercury surrounded by atmospheric air. Mercury droplets form spherical balls that can be rolled like a solid ball on a surface without wetting the surface. The surface tension of a liquid, in general, decreases with temperature and becomes zero at the critical point (and thus there is no distinct liquid–vapor interface at temperatures above the critical point). The effect of pressure on surface tension is usually negligible. The surface tension of a substance can be changed considerably by impurities. Therefore, certain chemicals, called surfactants, can be added to a liquid to decrease its surface tension. For example, soaps and detergents lower the surface tension of water and enable it to penetrate through the small openings between fibers for more effective washing. But this also means that devices whose operation depends on surface tension (such as heat pipes) can be destroyed by the presence of impurities due to poor workmanship. We speak of surface tension for liquids only at liquid–liquid or liquid–gas interfaces. Therefore, it is important to specify the adjacent liquid or gas when specifying surface tension. Also, surface tension determines the size of the liquid droplets that form. A droplet that keeps growing by the addition of more mass will break down when the surface tension can no longer hold it together. This is like a balloon that will burst while being inflated when the pressure inside rises above the strength of the balloon material. A curved interface indicates a pressure difference (or “pressure jump”) across the interface with pressure being higher on the concave side. The
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53 CHAPTER 2
excess pressure &P inside a droplet or bubble above the atmospheric pressure, for example, can be determined by considering the free-body diagram of half a droplet or bubble (Fig. 2–22). Noting that surface tension acts along the circumference and the pressure acts on the area, horizontal force balances for the droplet and the bubble give Droplet:
(2pR)ss ! (pR2)&Pdroplet → &Pdroplet ! Pi # Po !
2ss R
(2–36)
Bubble:
2(2pR)ss ! (pR2)&Pbubble → &Pbubble ! Pi # Po !
4ss R
(2–37)
where Pi and Po are the pressures inside and outside the droplet or bubble, respectively. When the droplet or bubble is in the atmosphere, Po is simply atmospheric pressure. The factor 2 in the force balance for the bubble is due to the bubble consisting of a film with two surfaces (inner and outer surfaces) and thus two circumferences in the cross section. The excess pressure in a droplet (or bubble) also can be determined by considering a differential increase in the radius of the droplet due to the addition of a differential amount of mass and interpreting the surface tension as the increase in the surface energy per unit area. Then the increase in the surface energy of the droplet during this differential expansion process becomes
(2π R)σs
(π R2)∆Pdroplet
(a) Half a droplet 2(2π R)σs
(π R2)∆Pbubble
(b) Half a bubble
FIGURE 2–22 The free-body diagram of half a droplet and half a bubble.
dWsurface ! ss dA ! ss d(4pR 2)! 8pRss dR
The expansion work done during this differential process is determined by multiplying the force by distance to obtain dWexpansion ! Force " Distance ! F dR ! (&PA) dR ! 4pR2 &P dR
Equating the two expressions above gives &Pdroplet ! 2ss /R, which is the same relation obtained before and given in Eq. 2–36. Note that the excess pressure in a droplet or bubble is inversely proportional to the radius.
Capillary Effect
Another interesting consequence of surface tension is the capillary effect, which is the rise or fall of a liquid in a small-diameter tube inserted into the liquid. Such narrow tubes or confined flow channels are called capillaries. The rise of kerosene through a cotton wick inserted into the reservoir of a kerosene lamp is due to this effect. The capillary effect is also partially responsible for the rise of water to the top of tall trees. The curved free surface of a liquid in a capillary tube is called the meniscus. It is commonly observed that water in a glass container curves up slightly at the edges where it touches the glass surface; but the opposite occurs for mercury: it curves down at the edges (Fig. 2–23). This effect is usually expressed by saying that water wets the glass (by sticking to it) while mercury does not. The strength of the capillary effect is quantified by the contact (or wetting) angle f, defined as the angle that the tangent to the liquid surface makes with the solid surface at the point of contact. The surface tension force acts along this tangent line toward the solid surface. A liquid is said to wet the surface when f + 90° and not to wet the surface when f 90°. In atmospheric air, the contact angle of water (and most other organic
f f Water
Mercury
(a) Wetting fluid
(b) Nonwetting fluid
FIGURE 2–23 The contact angle for wetting and nonwetting fluids.
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54 FLUID MECHANICS
FIGURE 2–24 The meniscus of colored water in a 4-mm-inner-diameter glass tube. Note that the edge of the meniscus meets the wall of the capillary tube at a very small contact angle. Photo by Gabrielle Trembley, Pennsylvania State University. Used by permission.
Meniscus
h>0
Meniscus h 0 and sin u > 0
m d du dP ar b ! ' rg sin u r dr dr dx
(8–32)
Downhill flow: u < 0 and sin u < 0
Following the same solution procedure, the velocity profile can be shown to be u(r) ! &
r2 R2 dP a ' rg sin ub a1 & 2b 4m dx R
(8–33)
It can also be shown that the average velocity and the volume flow rate relations for laminar flow through inclined pipes are, respectively, Vavg !
((P & rgL sin u)D2 32mL
and
# ((P & rgL sin u)pD4 V! 128mL
(8–34)
which are identical to the corresponding relations for horizontal pipes, except that (P is replaced by (P & rgL sin u. Therefore, the results already obtained for horizontal pipes can also be used for inclined pipes provided that (P is replaced by (P & rgL sin u (Fig. 8–16). Note that u ) 0 and thus sin u ) 0 for uphill flow, and u * 0 and thus sin u * 0 for downhill flow.
FIGURE 8–16 The relations developed for fully developed laminar flow through horizontal pipes can also be used for inclined pipes by replacing (P with (P & rgL sin u.
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332 FLUID MECHANICS
In inclined pipes, the combined effect of pressure difference and gravity drives the flow. Gravity helps downhill flow but opposes uphill flow. Therefore, much greater pressure differences need to be applied to maintain a specified flow rate in uphill flow although this becomes important only for liquids, because the density of gases is generally low. In the special case of . no flow (V ! 0), we have (P ! rgL sin u, which is what we would obtain from fluid statics (Chap. 3).
Laminar Flow in Noncircular Pipes
The friction factor f relations are given in Table 8–1 for fully developed laminar flow in pipes of various cross sections. The Reynolds number for flow in these pipes is based on the hydraulic diameter Dh ! 4Ac /p, where Ac is the cross-sectional area of the pipe and p is its wetted perimeter.
TA B L E 8 – 1 Friction factor for fully developed laminar flow in pipes of various cross sections (Dh ! 4Ac /p and Re ! Vavg Dh /n) Tube Geometry Circle
a/b or u°
Friction Factor f
—
64.00/Re
D
Rectangle
b a
Ellipse
b a
Isosceles triangle
u
a/b 1 2 3 4 6 8
+
56.92/Re 62.20/Re 68.36/Re 72.92/Re 78.80/Re 82.32/Re 96.00/Re
a/b 1 2 4 8 16
64.00/Re 67.28/Re 72.96/Re 76.60/Re 78.16/Re
u 10° 30° 60° 90° 120°
50.80/Re 52.28/Re 53.32/Re 52.60/Re 50.96/Re
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333 CHAPTER 8
EXAMPLE 8–1
Horizontal
Flow Rates in Horizontal and Inclined Pipes
Oil at 20°C (r ! 888 kg/m3 and m ! 0.800 kg/m · s) is flowing steadily through a 5-cm-diameter 40-m-long pipe (Fig. 8–17). The pressure at the pipe inlet and outlet are measured to be 745 and 97 kPa, respectively. Determine the flow rate of oil through the pipe assuming the pipe is (a) horizontal, (b) inclined 15° upward, (c) inclined 15° downward. Also verify that the flow through the pipe is laminar.
SOLUTION The pressure readings at the inlet and outlet of a pipe are given. The flow rates are to be determined for three different orientations, and the flow is to be shown to be laminar. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as a pump or a turbine. Properties The density and dynamic viscosity of oil are given to be r ! 888 kg/m3 and m ! 0.800 kg/m · s, respectively. Analysis The pressure drop across the pipe and the pipe cross-sectional area are
+15˚
–15˚
(P ! P1 & P2 ! 745 & 97 ! 648 kPa A c ! pD2/4 ! p(0.05 m)2/4 ! 0.001963 m2 (a) The flow rate for all three cases can be determined from Eq. 8–34,
# ((P & rgL sin u)pD4 V! 128mL where u is the angle the pipe makes with the horizontal. For the horizontal case, u ! 0 and thus sin u ! 0. Therefore,
# (648 kPa)p (0.05 m)4 1000 N/m2 1 kg , m/s2 (P pD4 ! a ba b V horiz ! 128mL 128(0.800 kg/m , s)(40 m) 1 kPa 1N ! 0.00311 m3/s
(b) For uphill flow with an inclination of 15°, we have u ! '15°, and
# ((P & rgL sin u)pD4 V uphill ! 128mL !
[648,000 Pa & (888 kg/m3)(9.81 m/s2)(40 m) sin 15°]p(0.05 m)4 1 kg , m/s2 a b 128(0.800 kg/m , s)(40 m) 1 Pa , m2
! 0.00267 m3/s
(c) For downhill flow with an inclination of 15°, we have u ! &15°, and
# ((P & rgL sin u)pD4 V downhill ! 128mL !
[648,000 Pa & (888 kg/m3)(9.81 m/s2)(40 m) sin (&15-)]p(0.05 m)4 1 kg , m/s2 a b 128(0.800 kg/m , s)(40 m) 1 Pa , m2
! 0.00354 m3/s
FIGURE 8–17 Schematic for Example 8–1.
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334 FLUID MECHANICS
The flow rate is the highest for the downhill flow case, as expected. The average fluid velocity and the Reynolds number in this case are
Vavg ! Re !
# V 0.00354 m3/s ! ! 1.80 m/s Ac 0.001963 m2 rVavgD m
!
(888 kg/m3)(1.80 m/s)(0.05 m) ! 100 0.800 kg/m , s
which is much less than 2300. Therefore, the flow is laminar for all three cases and the analysis is valid. Discussion Note that the flow is driven by the combined effect of pressure difference and gravity. As can be seen from the flow rates we calculated, gravity opposes uphill flow, but enhances downhill flow. Gravity has no effect on the flow rate in the horizontal case. Downhill flow can occur even in the absence of an applied pressure difference. For the case of P1 ! P2 ! 97 kPa (i.e., no applied pressure difference), the pressure throughout the entire pipe would remain constant at 97 Pa, and the fluid would flow through the pipe at a rate of 0.00043 m3/s under the influence of gravity. The flow rate increases as the tilt angle of the pipe from the horizontal is increased in the negative direction and would reach its maximum value when the pipe is vertical.
EXAMPLE 8–2
3.0 ft/s 0.12 in 30 ft
FIGURE 8–18 Schematic for Example 8–2.
Pressure Drop and Head Loss in a Pipe
Water at 40°F (r ! 62.42 lbm/ft3 and m ! 1.038 . 10&3 lbm/ft · s) is flowing through a 0.12-in- (! 0.010 ft) diameter 30-ft-long horizontal pipe steadily at an average velocity of 3.0 ft/s (Fig. 8–18). Determine (a) the head loss, (b) the pressure drop, and (c) the pumping power requirement to overcome this pressure drop.
SOLUTION The average flow velocity in a pipe is given. The head loss, the pressure drop, and the pumping power are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. Properties The density and dynamic viscosity of water are given to be r ! 62.42 lbm/ft3 and m ! 1.038 . 10&3 lbm/ft · s, respectively. Analysis (a) First we need to determine the flow regime. The Reynolds number is Re !
rVavgD m
!
(62.42 lbm/ft3)(3 ft/s)(0.01 ft) ! 1803 1.038 . 10 &3 lbm/ft , s
which is less than 2300. Therefore, the flow is laminar. Then the friction factor and the head loss become
f!
64 64 ! ! 0.0355 Re 1803
hL ! f
2 (3 ft/s)2 L V avg 30 ft ! 0.0355 ! 14.9 ft D 2g 0.01 ft 2(32.2 ft/s2)
(b) Noting that the pipe is horizontal and its diameter is constant, the pressure drop in the pipe is due entirely to the frictional losses and is equivalent to the pressure loss,
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335 CHAPTER 8
(P ! (PL ! f
2 L rV avg 30 ft (62.42 lbm/ft3)(3 ft/s)2 1 lbf ! 0.0355 a b D 2 0.01 ft 2 32.2 lbm , ft/s2
! 929 lbf/ft2 ! 6.45 psi
(c) The volume flow rate and the pumping power requirements are
# V ! Vavg A c ! Vavg(pD2/4) ! (3 ft/s)[p(0.01 ft)2/4] ! 0.000236 ft3/s
# # 1W W pump ! V (P ! (0.000236 ft3/s)(929 lbf/ft2) a b ! 0.30 W 0.737 lbf , ft/s
Therefore, power input in the amount of 0.30 W is needed to overcome the frictional losses in the flow due to viscosity. Discussion The pressure rise provided by a pump is often listed by a pump manufacturer in units of head (Chap. 14). Thus, the pump in this flow needs to provide 14.9 ft of water head in order to overcome the irreversible head loss.
8–5
■
TURBULENT FLOW IN PIPES
Most flows encountered in engineering practice are turbulent, and thus it is important to understand how turbulence affects wall shear stress. However, turbulent flow is a complex mechanism dominated by fluctuations, and despite tremendous amounts of work done in this area by researchers, the theory of turbulent flow remains largely undeveloped. Therefore, we must rely on experiments and the empirical or semi-empirical correlations developed for various situations. Turbulent flow is characterized by random and rapid fluctuations of swirling regions of fluid, called eddies, throughout the flow. These fluctuations provide an additional mechanism for momentum and energy transfer. In laminar flow, fluid particles flow in an orderly manner along pathlines, and momentum and energy are transferred across streamlines by molecular diffusion. In turbulent flow, the swirling eddies transport mass, momentum, and energy to other regions of flow much more rapidly than molecular diffusion, greatly enhancing mass, momentum, and heat transfer. As a result, turbulent flow is associated with much higher values of friction, heat transfer, and mass transfer coefficients (Fig. 8–19). Even when the average flow is steady, the eddy motion in turbulent flow causes significant fluctuations in the values of velocity, temperature, pressure, and even density (in compressible flow). Figure 8–20 shows the variation of the instantaneous velocity component u with time at a specified location, as can be measured with a hot-wire anemometer probe or other sensitive device. We observe that the instantaneous values of the velocity fluctuate about an average value, which suggests that the velocity can be expressed as the sum of an average value u– and a fluctuating component u/, u ! u ' u/
2 2 2 2 2 5 5 5 5 5 7 7 7 7 7 12 12 12 12 12 (a) Before turbulence
12 2 5 7 5 2 5 7 2 12 7 12 7 5 12 2 7 5 12 2 (b) After turbulence
FIGURE 8–19 The intense mixing in turbulent flow brings fluid particles at different momentums into close contact and thus enhances momentum transfer.
u
u–
u/
(8–35)
This is also the case for other properties such as the velocity component v – – in the y-direction, and thus v ! –v ' v/, P ! P ' P/, and T ! T ' T/. The average value of a property at some location is determined by averaging it over a time interval that is sufficiently large so that the time average levels off to a constant. Therefore, the time average of fluctuating components is
Time, t
FIGURE 8–20 Fluctuations of the velocity component u with time at a specified location in turbulent flow.
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336 FLUID MECHANICS
zero, e.g., u/ ! 0. The magnitude of u/ is usually just a few percent of u–, but the high frequencies of eddies (in the order of a thousand per second) makes them very effective for the transport of momentum, thermal energy, and mass. In time-averaged stationary turbulent flow, the average values of properties (indicated by an overbar) are independent of time. The chaotic fluctuations of fluid particles play a dominant role in pressure drop, and these random motions must be considered in analyses together with the average velocity. Perhaps the first thought that comes to mind is to determine the shear stress in an analogous manner to laminar flow from t ! &m du–/dr, where u–(r) is the average velocity profile for turbulent flow. But the experimental studies show that this is not the case, and the shear stress is much larger due to the turbulent fluctuations. Therefore, it is convenient to think of the turbulent shear stress as consisting of two parts: the laminar component, which accounts for the friction between layers in the flow direction (expressed as tlam ! &m du–/dr), and the turbulent component, which accounts for the friction between the fluctuating fluid particles and the fluid body (denoted as tturb and is related to the fluctuation components of velocity). Then the total shear stress in turbulent flow can be expressed as
r
ttotal ! tlam ' tturb
0 u(r)
r 0
ttotal
0 tlam
t tturb
FIGURE 8–21 The velocity profile and the variation of shear stress with radial distance for turbulent flow in a pipe.
y
rv / dA dA
u(y)
v/ u/
u
FIGURE 8–22 Fluid particle moving upward through a differential area dA as a result of the velocity fluctuation v/.
(8–36)
The typical average velocity profile and relative magnitudes of laminar and turbulent components of shear stress for turbulent flow in a pipe are given in Fig. 8–21. Note that although the velocity profile is approximately parabolic in laminar flow, it becomes flatter or “fuller” in turbulent flow, with a sharp drop near the pipe wall. The fullness increases with the Reynolds number, and the velocity profile becomes more nearly uniform, lending support to the commonly utilized uniform velocity profile approximation for fully developed turbulent pipe flow. Keep in mind, however, that the flow speed at the wall of a stationary pipe is always zero (no-slip condition).
Turbulent Shear Stress
Consider turbulent flow in a horizontal pipe, and the upward eddy motion of fluid particles in a layer of lower velocity to an adjacent layer of higher velocity through a differential area dA as a result of the velocity fluctuation v/, as shown in Fig. 8–22. The mass flow rate of the fluid particles rising through dA is rv/dA, and its net effect on the layer above dA is a reduction in its average flow velocity because of momentum transfer to the fluid particles with lower average flow velocity. This momentum transfer causes the horizontal velocity of the fluid particles to increase by u/, and thus its momentum in the horizontal direction to increase at a rate of (rv/dA)u/, which must be equal to the decrease in the momentum of the upper fluid layer. Noting that force in a given direction is equal to the rate of change of momentum in that direction, the horizontal force acting on a fluid element above dA due to the passing of fluid particles through dA is dF ! (rv/dA)(&u/) ! &ru/v/dA. Therefore, the shear force per unit area due to the eddy motion of fluid particles dF/dA ! &ru/v/ can be viewed as the instantaneous turbulent shear stress. Then the turbulent shear stress can be expressed as tturb ! &ru/v/
(8–37)
where u/v/ is the time average of the product of the fluctuating velocity components u/ and v/. Note that u/v/ 0 0 even though u/ ! 0 and v/ ! 0
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337 CHAPTER 8
(and thus u/ v/ ! 0), and experimental results show that u/v/ is usually a negative quantity. Terms such as &ru/v/ or &ru/2 are called Reynolds stresses or turbulent stresses. Many semi-empirical formulations have been developed that model the Reynolds stress in terms of average velocity gradients in order to provide mathematical closure to the equations of motion. Such models are called turbulence models and are discussed in more detail in Chap. 15. The random eddy motion of groups of particles resembles the random motion of molecules in a gas—colliding with each other after traveling a certain distance and exchanging momentum in the process. Therefore, momentum transport by eddies in turbulent flows is analogous to the molecular momentum diffusion. In many of the simpler turbulence models, turbulent shear stress is expressed in an analogous manner as suggested by the French mathematician Joseph Boussinesq (1842–1929) in 1877 as tturb ! &ru/v/ ! m t
%u %y
y
(8–38)
where mt is the eddy viscosity or turbulent viscosity, which accounts for momentum transport by turbulent eddies. Then the total shear stress can be expressed conveniently as %u %u ttotal ! (m ' m t) ! r(n ' nt) %y %y
%u %u 2 ! rl 2m a b %y %y
%u %y
b
y=0
(8–39)
where nt ! mt /r is the kinematic eddy viscosity or kinematic turbulent viscosity (also called the eddy diffusivity of momentum). The concept of eddy viscosity is very appealing, but it is of no practical use unless its value can be determined. In other words, eddy viscosity must be modeled as a function of the average flow variables; we call this eddy viscosity closure. For example, in the early 1900s, the German engineer L. Prandtl introduced the concept of mixing length lm, which is related to the average size of the eddies that are primarily responsible for mixing, and expressed the turbulent shear stress as tturb ! m t
Laminar flow
a
y
(8–40)
But this concept is also of limited use since lm is not a constant for a given flow (in the vicinity of the wall, for example, lm is nearly proportional to the distance from the wall) and its determination is not easy. Final mathematical closure is obtained only when lm is written as a function of average flow variables, distance from the wall, etc. Eddy motion and thus eddy diffusivities are much larger than their molecular counterparts in the core region of a turbulent boundary layer. The eddy motion loses its intensity close to the wall and diminishes at the wall because of the no-slip condition (u/ and v/ are identically zero at a stationary wall). Therefore, the velocity profile is very slowly changing in the core region of a turbulent boundary layer, but very steep in the thin layer adjacent to the wall, resulting in large velocity gradients at the wall surface. So it is no surprise that the wall shear stress is much larger in turbulent flow than it is in laminar flow (Fig. 8–23). Note that molecular diffusivity of momentum n (as well as m) is a fluid property, and its value is listed in fluid handbooks. Eddy diffusivity nt (as well as mt), however, is not a fluid property, and its value depends on flow
a
%u %y
b
y=0
Turbulent flow
FIGURE 8–23 The velocity gradients at the wall, and thus the wall shear stress, are much larger for turbulent flow than they are for laminar flow, even though the turbulent boundary layer is thicker than the laminar one for the same value of free-stream velocity.
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338 FLUID MECHANICS
conditions. Eddy diffusivity nt decreases toward the wall, becoming zero at the wall. Its value ranges from zero at the wall to several thousand times the value of the molecular diffusivity in the core region.
Vavg u(r)
r
Turbulent Velocity Profile
0
Laminar flow
Vavg r
u(r) 0
Turbulent layer Overlap layer Turbulent flow
Buffer layer Viscous sublayer
FIGURE 8–24 The velocity profile in fully developed pipe flow is parabolic in laminar flow, but much fuller in turbulent flow.
Unlike laminar flow, the expressions for the velocity profile in a turbulent flow are based on both analysis and measurements, and thus they are semi-empirical in nature with constants determined from experimental data. Consider fully developed turbulent flow in a pipe, and let u denote the timeaveraged velocity in the axial direction (and thus drop the overbar from u– for simplicity). Typical velocity profiles for fully developed laminar and turbulent flows are given in Fig. 8–24. Note that the velocity profile is parabolic in laminar flow but is much fuller in turbulent flow, with a sharp drop near the pipe wall. Turbulent flow along a wall can be considered to consist of four regions, characterized by the distance from the wall. The very thin layer next to the wall where viscous effects are dominant is the viscous (or laminar or linear or wall) sublayer. The velocity profile in this layer is very nearly linear, and the flow is streamlined. Next to the viscous sublayer is the buffer layer, in which turbulent effects are becoming significant, but the flow is still dominated by viscous effects. Above the buffer layer is the overlap (or transition) layer, also called the inertial sublayer, in which the turbulent effects are much more significant, but still not dominant. Above that is the outer (or turbulent) layer in the remaining part of the flow in which turbulent effects dominate over molecular diffusion (viscous) effects. Flow characteristics are quite different in different regions, and thus it is difficult to come up with an analytic relation for the velocity profile for the entire flow as we did for laminar flow. The best approach in the turbulent case turns out to be to identify the key variables and functional forms using dimensional analysis, and then to use experimental data to determine the numerical values of any constants. The thickness of the viscous sublayer is very small (typically, much less than 1 percent of the pipe diameter), but this thin layer next to the wall plays a dominant role on flow characteristics because of the large velocity gradients it involves. The wall dampens any eddy motion, and thus the flow in this layer is essentially laminar and the shear stress consists of laminar shear stress which is proportional to the fluid viscosity. Considering that velocity changes from zero to nearly the core region value across a layer that is sometimes no thicker than a hair (almost like a step function), we would expect the velocity profile in this layer to be very nearly linear, and experiments confirm that. Then the velocity gradient in the viscous sublayer remains nearly constant at du/dy ! u/y, and the wall shear stress can be expressed as u u tw ! m ! rn y y
or
tw nu ! r y
(8–41)
where y is the distance from the wall (note that y ! R & r for a circular pipe). The quantity tw /r is frequently encountered in the analysis of turbulent velocity profiles. The square root of tw /r has the dimensions of velocity, and thus it is convenient to view it as a fictitious velocity called the friction velocity expressed as u * ! 1tw /r. Substituting this into Eq. 8–41, the velocity profile in the viscous sublayer can be expressed in dimensionless form as
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339 CHAPTER 8
yu * u ! n u*
Viscous sublayer:
(8–42)
This equation is known as the law of the wall, and it is found to satisfactorily correlate with experimental data for smooth surfaces for 0 $ yu*/n $ 5. Therefore, the thickness of the viscous sublayer is roughly Thickness of viscous sublayer:
y ! d sublayer !
5n 25n ! u* ud
(8–43)
where ud is the flow velocity at the edge of the viscous sublayer, which is closely related to the average velocity in a pipe. Thus we conclude that the thickness of the viscous sublayer is proportional to the kinematic viscosity and inversely proportional to the average flow velocity. In other words, the viscous sublayer is suppressed and it gets thinner as the velocity (and thus the Reynolds number) increases. Consequently, the velocity profile becomes nearly flat and thus the velocity distribution becomes more uniform at very high Reynolds numbers. The quantity n/u* has dimensions of length and is called the viscous length; it is used to nondimensionalize the distance y from the surface. In boundary layer analysis, it is convenient to work with nondimensionalized distance and nondimensionalized velocity defined as Nondimensionalized variables:
y' !
yu * n
and
u' !
u u*
(8–44)
Then the law of the wall (Eq. 8–42) becomes simply Normalized law of the wall:
u' ! y '
(8–45)
Note that the friction velocity u* is used to nondimensionalize both y and u, and y' resembles the Reynolds number expression. In the overlap layer, the experimental data for velocity are observed to line up on a straight line when plotted against the logarithm of distance from the wall. Dimensional analysis indicates and the experiments confirm that the velocity in the overlap layer is proportional to the logarithm of distance, and the velocity profile can be expressed as The logarithmic law:
1 yu * u 'B ! ln n u* k
(8–46)
where k and B are constants whose values are determined experimentally to be about 0.40 and 5.0, respectively. Equation 8–46 is known as the logarithmic law. Substituting the values of the constants, the velocity profile is determined to be Overlap layer:
yu * u ' 5.0 ! 2.5 ln n u*
or
u ' ! 2.5 ln y ' ' 5.0
(8–47)
It turns out that the logarithmic law in Eq. 8–47 satisfactorily represents experimental data for the entire flow region except for the regions very close to the wall and near the pipe center, as shown in Fig. 8–25, and thus it is viewed as a universal velocity profile for turbulent flow in pipes or over surfaces. Note from the figure that the logarithmic-law velocity profile is quite accurate for y' ) 30, but neither velocity profile is accurate in the buffer layer, i.e., the region 5 * y' * 30. Also, the viscous sublayer appears much larger in the figure than it is since we used a logarithmic scale for distance from the wall.
u+ = u/u* 30 25 20 15
Eq. 8–42 Eq. 8–47
10 Experimental data
5 0 0 10
101
Viscous Buffer sublayer layer
102 y+ = yu*/n Overlap layer
103
104
Turbulent layer
FIGURE 8–25 Comparison of the law of the wall and the logarithmic-law velocity profiles with experimental data for fully developed turbulent flow in a pipe.
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340 FLUID MECHANICS
A good approximation for the outer turbulent layer of pipe flow can be obtained by evaluating the constant B in Eq. 8–46 from the requirement that maximum velocity in a pipe occurs at the centerline where r ! 0. Solving for B from Eq. 8–46 by setting y ! R & r ! R and u ! umax, and substituting it back into Eq. 8–46 together with k ! 0.4 gives umax & u R ! 2.5 ln u* R&r
Outer turbulent layer:
1
n = 10 n=8
0.8
r/R
0.6
n=6 Laminar
0.2 0
0
0.2
0.4 0.6 u/umax
The deviation of velocity from the centerline value umax & u is called the velocity defect, and Eq. 8–48 is called the velocity defect law. This relation shows that the normalized velocity profile in the core region of turbulent flow in a pipe depends on the distance from the centerline and is independent of the viscosity of the fluid. This is not surprising since the eddy motion is dominant in this region, and the effect of fluid viscosity is negligible. Numerous other empirical velocity profiles exist for turbulent pipe flow. Among those, the simplest and the best known is the power-law velocity profile expressed as Power-law velocity profile:
0.4
0.8
FIGURE 8–26 Power-law velocity profiles for fully developed turbulent flow in a pipe for different exponents, and its comparison with the laminar velocity profile.
1
(8–48)
y 1/n !a b u max R u
or
r 1/n ! a1 & b u max R u
(8–49)
where the exponent n is a constant whose value depends on the Reynolds number. The value of n increases with increasing Reynolds number. The value n ! 7 generally approximates many flows in practice, giving rise to the term one-seventh power-law velocity profile. Various power-law velocity profiles are shown in Fig. 8–26 for n ! 6, 8, and 10 together with the velocity profile for fully developed laminar flow for comparison. Note that the turbulent velocity profile is fuller than the laminar one, and it becomes more flat as n (and thus the Reynolds number) increases. Also note that the power-law profile cannot be used to calculate wall shear stress since it gives a velocity gradient of infinity there, and it fails to give zero slope at the centerline. But these regions of discrepancy constitute a small portion of flow, and the power-law profile gives highly accurate results for turbulent flow through a pipe. Despite the small thickness of the viscous sublayer (usually much less than 1 percent of the pipe diameter), the characteristics of the flow in this layer are very important since they set the stage for flow in the rest of the pipe. Any irregularity or roughness on the surface disturbs this layer and affects the flow. Therefore, unlike laminar flow, the friction factor in turbulent flow is a strong function of surface roughness. It should be kept in mind that roughness is a relative concept, and it has significance when its height e is comparable to the thickness of the laminar sublayer (which is a function of the Reynolds number). All materials appear “rough” under a microscope with sufficient magnification. In fluid mechanics, a surface is characterized as being rough when the hills of roughness protrude out of the laminar sublayer. A surface is said to be smooth when the sublayer submerges the roughness elements. Glass and plastic surfaces are generally considered to be hydrodynamically smooth.
The Moody Chart
The friction factor in fully developed turbulent pipe flow depends on the Reynolds number and the relative roughness e/D, which is the ratio of the
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341 CHAPTER 8
mean height of roughness of the pipe to the pipe diameter. The functional form of this dependence cannot be obtained from a theoretical analysis, and all available results are obtained from painstaking experiments using artificially roughened surfaces (usually by gluing sand grains of a known size on the inner surfaces of the pipes). Most such experiments were conducted by Prandtl’s student J. Nikuradse in 1933, followed by the works of others. The friction factor was calculated from the measurements of the flow rate and the pressure drop. The experimental results obtained are presented in tabular, graphical, and functional forms obtained by curve-fitting experimental data. In 1939, Cyril F. Colebrook (1910–1997) combined the available data for transition and turbulent flow in smooth as well as rough pipes into the following implicit relation known as the Colebrook equation: 1 2f
! &2.0 loga
e/D 2.51 b ' 3.7 Re 2f
(turbulent flow)
(8–50)
We note that the logarithm in Eq. 8–50 is a base 10 rather than a natural logarithm. In 1942, the American engineer Hunter Rouse (1906–1996) verified Colebrook’s equation and produced a graphical plot of f as a function of Re and the product Re1f . He also presented the laminar flow relation and a table of commercial pipe roughness. Two years later, Lewis F. Moody (1880–1953) redrew Rouse’s diagram into the form commonly used today. The now famous Moody chart is given in the appendix as Fig. A–12. It presents the Darcy friction factor for pipe flow as a function of the Reynolds number and e/D over a wide range. It is probably one of the most widely accepted and used charts in engineering. Although it is developed for circular pipes, it can also be used for noncircular pipes by replacing the diameter by the hydraulic diameter. Commercially available pipes differ from those used in the experiments in that the roughness of pipes in the market is not uniform and it is difficult to give a precise description of it. Equivalent roughness values for some commercial pipes are given in Table 8–2 as well as on the Moody chart. But it should be kept in mind that these values are for new pipes, and the relative roughness of pipes may increase with use as a result of corrosion, scale buildup, and precipitation. As a result, the friction factor may increase by a factor of 5 to 10. Actual operating conditions must be considered in the design of piping systems. Also, the Moody chart and its equivalent Colebrook equation involve several uncertainties (the roughness size, experimental error, curve fitting of data, etc.), and thus the results obtained should not be treated as “exact.” It is usually considered to be accurate to 115 percent over the entire range in the figure. The Colebrook equation is implicit in f, and thus the determination of the friction factor requires some iteration unless an equation solver such as EES is used. An approximate explicit relation for f was given by S. E. Haaland in 1983 as 1 2f
e/D 1.11 6.9 'a b d Re 3.7
" &1.8 logc
(8–51)
The results obtained from this relation are within 2 percent of those obtained from the Colebrook equation. If more accurate results are desired, Eq. 8–51 can be used as a good first guess in a Newton iteration when using a programmable calculator or a spreadsheet to solve for f with Eq. 8–50.
TA B L E 8 – 2 Equivalent roughness values for new commercial pipes* Roughness, e Material Glass, plastic Concrete Wood stave Rubber, smoothed Copper or brass tubing Cast iron Galvanized iron Wrought iron Stainless steel Commercial steel
ft
mm
0 (smooth) 0.003–0.03 0.9–9 0.0016 0.5 0.000033
0.01
0.000005 0.00085
0.0015 0.26
0.0005 0.00015 0.000007
0.15 0.046 0.002
0.00015
0.045
* The uncertainty in these values can be as much as 160 percent.
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342 FLUID MECHANICS
Relative Roughness, 2/D
Friction Factor, f
0.0* 0.00001 0.0001 0.0005 0.001 0.005 0.01 0.05
0.0119 0.0119 0.0134 0.0172 0.0199 0.0305 0.0380 0.0716
* Smooth surface. All values are for Re ! 106 and are calculated from the Colebrook equation.
FIGURE 8–27 The friction factor is minimum for a smooth pipe and increases with roughness.
We make the following observations from the Moody chart: • For laminar flow, the friction factor decreases with increasing Reynolds number, and it is independent of surface roughness. • The friction factor is a minimum for a smooth pipe (but still not zero because of the no-slip condition) and increases with roughness (Fig. 8–27). The Colebrook equation in this case (e ! 0) reduces to the Prandtl equation expressed as 1/ 1f ! 2.0 log(Re 1f ) & 0.8. • The transition region from the laminar to turbulent regime (2300 * Re * 4000) is indicated by the shaded area in the Moody chart (Figs. 8–28 and A–12). The flow in this region may be laminar or turbulent, depending on flow disturbances, or it may alternate between laminar and turbulent, and thus the friction factor may also alternate between the values for laminar and turbulent flow. The data in this range are the least reliable. At small relative roughnesses, the friction factor increases in the transition region and approaches the value for smooth pipes. • At very large Reynolds numbers (to the right of the dashed line on the chart) the friction factor curves corresponding to specified relative roughness curves are nearly horizontal, and thus the friction factors are independent of the Reynolds number (Fig. 8–28). The flow in that region is called fully rough turbulent flow or just fully rough flow because the thickness of the viscous sublayer decreases with increasing Reynolds number, and it becomes so thin that it is negligibly small compared to the surface roughness height. The viscous effects in this case are produced in the main flow primarily by the protruding roughness elements, and the contribution of the laminar sublayer is negligible. The Colebrook equation in the fully rough zone (Re → +) reduces to the von Kármán equation expressed as 1/ 1f ! &2.0 log[(e/D)/3.7], which is explicit in f. Some authors call this zone completely (or fully) turbulent flow, but this is misleading since the flow to the left of the dashed blue line in Fig. 8–28 is also fully turbulent. In calculations, we should make sure that we use the actual internal diameter of the pipe, which may be different than the nominal diameter. For example, the internal diameter of a steel pipe whose nominal diameter is 1 in is 1.049 in (Table 8–3). 0.1
Laminar
Fully rough turbulent flow (ƒ levels off) e /D = 0.01 e /D = 0.001
ƒ
Transitional
e /D = 0.0001
0.01
FIGURE 8–28 At very large Reynolds numbers, the friction factor curves on the Moody chart are nearly horizontal, and thus the friction factors are independent of the Reynolds number.
e /D = 0
Smooth turbulent
0.001 103
104
105
106 Re
107
108
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343 CHAPTER 8
Types of Fluid Flow Problems
In the design and analysis of piping systems that involve the use of the Moody chart (or the Colebrook equation), we usually encounter three types of problems (the fluid and the roughness of the pipe are assumed to be specified in all cases) (Fig. 8–29): Determining the pressure drop (or head loss) when the pipe length and diameter are given for a specified flow rate (or velocity) 2. Determining the flow rate when the pipe length and diameter are given for a specified pressure drop (or head loss) 3. Determining the pipe diameter when the pipe length and flow rate are given for a specified pressure drop (or head loss)
TA B L E 8 – 3 Standard sizes for Schedule 40 steel pipes Nominal Size, in
1.
Problems of the first type are straightforward and can be solved directly by using the Moody chart. Problems of the second type and third type are commonly encountered in engineering design (in the selection of pipe diameter, for example, that minimizes the sum of the construction and pumping costs), but the use of the Moody chart with such problems requires an iterative approach unless an equation solver is used. In problems of the second type, the diameter is given but the flow rate is unknown. A good guess for the friction factor in that case is obtained from the completely turbulent flow region for the given roughness. This is true for large Reynolds numbers, which is often the case in practice. Once the flow rate is obtained, the friction factor can be corrected using the Moody chart or the Colebrook equation, and the process is repeated until the solution converges. (Typically only a few iterations are required for convergence to three or four digits of precision.) In problems of the third type, the diameter is not known and thus the Reynolds number and the relative roughness cannot be calculated. Therefore, we start calculations by assuming a pipe diameter. The pressure drop calculated for the assumed diameter is then compared to the specified pressure drop, and calculations are repeated with another pipe diameter in an iterative fashion until convergence. To avoid tedious iterations in head loss, flow rate, and diameter calculations, Swamee and Jain proposed the following explicit relations in 1976 that are accurate to within 2 percent of the Moody chart: # 10 &6 * e/D * 10 &2 nD 0.9 &2 V 2L e ' 4.62a # b df hL ! 1.07 5 eln c 3.7D 3000 * Re * 3 . 10 8 gD V
# gD5hL 0.5 3.17v 2L 0.5 e b ln c 'a V ! &0.965a b d Re ) 2000 L 3.7D gD3h L # # LV 2 4.75 L 5.2 0.04 10 &6 * e/D * 10 &2 D ! 0.66ce 1.25 a b ' nV 9.4 a b d ghL ghL 5000 * Re * 3 . 10 8
(8–52)
(8–53)
(8–54)
Note that all quantities are dimensional and the units simplify to the desired unit (for example, to m or ft in the last relation) when consistent units are used. Noting that the Moody chart is accurate to within 15 percent of experimental data, we should have no reservation in using these approximate relations in the design of piping systems.
Actual Inside Diameter, in
1 8 1 4 3 8 1 2 3 4
0.269 0.364 0.493 0.622 0.824 1.049 1.610 2.067 2.469 3.068 5.047 10.02
1 112 2 212 3 5 10
Problem type 1 2 3
Given ⋅ L,, D,, V L,, D,, ∆P ⋅ L,, ∆P,, V
Find ∆P (or hL) ⋅ V D
FIGURE 8–29 The three types of problems encountered in pipe flow.
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344 FLUID MECHANICS 0.2 ft3/s water
2 in
200 ft
FIGURE 8–30 Schematic for Example 8–3.
EXAMPLE 8–3
Determining the Head Loss in a Water Pipe
Water at 60°F (r ! 62.36 lbm/ft3 and m ! 7.536 . 10&4 lbm/ft · s) is flowing steadily in a 2-in-diameter horizontal pipe made of stainless steel at a rate of 0.2 ft3/s (Fig. 8–30). Determine the pressure drop, the head loss, and the required pumping power input for flow over a 200-ft-long section of the pipe.
SOLUTION The flow rate through a specified water pipe is given. The pressure drop, the head loss, and the pumping power requirements are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as a pump or a turbine. Properties The density and dynamic viscosity of water are given to be r ! 62.36 lbm/ft3 and m ! 7.536 . 10&4 lbm/ft · s, respectively. Analysis We recognize this as a problem of the first type, since flow rate, pipe length, and pipe diameter are known. First we calculate the average velocity and the Reynolds number to determine the flow regime: V ! Re !
# # V V 0.2 ft3/s ! ! ! 9.17 ft/s 2 Ac pD /4 p(2/12 ft)2/4
rV D (62.36 lbm/ft3)(9.17 ft/s)(2/12 ft) ! ! 126,400 m 7.536 . 10 &4 lbm/ft , s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is calculated using Table 8–2
e/D !
0.000007 ft ! 0.000042 2/12 ft
The friction factor corresponding to this relative roughness and the Reynolds number can simply be determined from the Moody chart. To avoid any reading error, we determine f from the Colebrook equation:
1 2f
e/D 2.51 1 0.000042 2.51 b→ ! &2.0 log a b ' ' 3.7 Re2f 3.7 2f 126,400 2f
! &2.0 loga
Using an equation solver or an iterative scheme, the friction factor is determined to be f ! 0.0174. Then the pressure drop (which is equivalent to pressure loss in this case), head loss, and the required power input become (P ! (PL ! f
L rV 2 200 ft (62.36 lbm/ft3)(9.17 ft/s)2 1 lbf ! 0.0174 a b D 2 2/12 ft 2 32.2 lbm , ft/s2
! 1700 lbf/ft2 ! 11.8 psi hL !
(PL L V2 200 ft (9.17 ft/s)2 ! f ! 0.0174 ! 27.3 ft rg D 2g 2/12 ft 2(32.2 ft/s2)
# # 1W W pump ! V (P ! (0.2 ft3/s)(1700 lbf/ft2)a b ! 461 W 0.737 lbf , ft/s
Therefore, power input in the amount of 461 W is needed to overcome the frictional losses in the pipe. Discussion It is common practice to write our final answers to three significant digits, even though we know that the results are accurate to at most two significant digits because of inherent inaccuracies in the Colebrook equation,
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345 CHAPTER 8
as discussed previously. The friction factor could also be determined easily from the explicit Haaland relation (Eq. 8–51). It would give f ! 0.0172, which is sufficiently close to 0.0174. Also, the friction factor corresponding to e ! 0 in this case is 0.0171, which indicates that stainless-steel pipes can be assumed to be smooth with negligible error.
EXAMPLE 8–4
Determining the Diameter of an Air Duct
Heated air at 1 atm and 35°C is to be transported in a 150-m-long circular plastic duct at a rate of 0.35 m3/s (Fig. 8–31). If the head loss in the pipe is not to exceed 20 m, determine the minimum diameter of the duct.
SOLUTION The flow rate and the head loss in an air duct are given. The diameter of the duct is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The duct involves no components such as bends, valves, and connectors. 4 Air is an ideal gas. 5 The duct is smooth since it is made of plastic. 6 The flow is turbulent (to be verified). Properties The density, dynamic viscosity, and kinematic viscosity of air at 35°C are r ! 1.145 kg/m3, m ! 1.895 . 10&5 kg/m · s, and n ! 1.655 . 10&5 m2/s. Analysis This is a problem of the third type since it involves the determination of diameter for specified flow rate and head loss. We can solve this problem by three different approaches: (1) an iterative approach by assuming a pipe diameter, calculating the head loss, comparing the result to the specified head loss, and repeating calculations until the calculated head loss matches the specified value; (2) writing all the relevant equations (leaving the diameter as an unknown) and solving them simultaneously using an equation solver; and (3) using the third Swamee–Jain formula. We will demonstrate the use of the last two approaches. The average velocity, the Reynolds number, the friction factor, and the head loss relations can be expressed as (D is in m, V is in m/s, and Re and f are dimensionless)
# # V V 0.35 m3/s ! ! 2 Ac pD /4 pD2/4 VD VD Re ! ! n 1.655 . 10 &5 m2/s e/D 2.51 2.51 1 ! &2.0 log a ' b ! &2.0 loga b 3.7 2f Re2f Re 2f L V2 150 m V2 → 20 ! f hL ! f D 2g D 2(9.81 m/s2) V !
The roughness is approximately zero for a plastic pipe (Table 8–2). Therefore, this is a set of four equations in four unknowns, and solving them with an equation solver such as EES gives
D ! 0.267 m,
f ! 0.0180,
V ! 6.24 m/s,
and
Re ! 100,800
Therefore, the diameter of the duct should be more than 26.7 cm if the head loss is not to exceed 20 m. Note that Re ) 4000, and thus the turbulent flow assumption is verified.
0.35 m3/s air
D
150 m
FIGURE 8–31 Schematic for Example 8–4
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346 FLUID MECHANICS
The diameter can also be determined directly from the third Swamee–Jain formula to be
# # LV 2 4.75 L 5.2 0.04 b ' nV 9.4 a b d gh L ghL
D ! 0.66ce 1.25 a
5.2 0.04 150 m d b (9.81 m/s2)(20 m)
! 0.66 c0 ' (1.655 . 10 &5 m2/s)(0.35 m3/s)9.4 a ! 0.271 m
Discussion Note that the difference between the two results is less than 2 percent. Therefore, the simple Swamee–Jain relation can be used with confidence. Finally, the first (iterative) approach requires an initial guess for D. If we use the Swamee–Jain result as our initial guess, the diameter converges to D ! 0.267 m in short order.
EXAMPLE 8–5
Determining the Flow Rate of Air in a Duct
Reconsider Example 8–4. Now the duct length is doubled while its diameter is maintained constant. If the total head loss is to remain constant, determine the drop in the flow rate through the duct.
SOLUTION The diameter and the head loss in an air duct are given. The drop in the flow rate is to be determined. Analysis This is a problem of the second type since it involves the determination of the flow rate for a specified pipe diameter and head loss. The solution involves an iterative approach since the flow rate (and thus the flow velocity) is not known. The average velocity, Reynolds number, friction factor, and the head loss relations can be expressed as (D is in m, V is in m/s, and Re and f are dimensionless) V! Re ! 1 2f
# # V V ! A c pD2/4
→
VD n
Re !
→
V!
V(0.267 m) 1.655 . 10 &5 m2/s
e/D 2.51 ' b 3.7 Re 2f
→
! &2.0 loga
hL ! f
L V2 D 2g
→
# V p(0.267 m)2/4
20 ! f
1
2.51 ! &2.0 log a b 2f Re 2f
300 m V2 0.267 m 2(9.81 m/s2)
This is a set of four equations in four unknowns and solving them with an equation solver such as EES gives
# V ! 0.24 m3/s,
f ! 0.0195,
V ! 4.23 m/s,
and
Re ! 68,300
Then the drop in the flow rate becomes
# # # V drop ! V old & V new ! 0.35 & 0.24 ! 0.11 m3/s
(a drop of 31 percent)
Therefore, for a specified head loss (or available head or fan pumping power), the flow rate drops by about 31 percent from 0.35 to 0.24 m3/s when the duct length doubles.
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347 CHAPTER 8
Alternative Solution If a computer is not available (as in an exam situation), another option is to set up a manual iteration loop. We have found that the best convergence is usually realized by first guessing the friction factor f, and then solving for the velocity V. The equation for V as a function of f is
Average velocity through the pipe:
V !
2ghL Bf L/D ˛
Now that V is calculated, the Reynolds number can be calculated, from which a corrected friction factor is obtained from the Moody chart or the Colebrook equation. We repeat the calculations with the corrected value of f until convergence. We guess f ! 0.04 for illustration: Iteration 1 2 3 4 5
f (guess)
V, m/s
0.04 0.0212 0.01973 0.01957 0.01956
2.955 4.059 4.207 4.224 4.225
Re 4.724 6.489 6.727 6.754 6.756
. . . . .
Corrected f 104 104 104 104 104
0.0212 0.01973 0.01957 0.01956 0.01956
Notice that the iteration has converged to three digits in only three iterations and to four digits in only four iterations. The final results are identical to those obtained with EES, yet do not require a computer. Discussion The new flow rate can also be determined directly from the second Swamee–Jain formula to be
# gD5h L 0.5 3.17v 2L 0.5 e b lnc 'a b d V ! &0.965a L 3.7D gD3hL ! &0.965a
(9.81 m/s2)(0.267 m)5(20 m) 0.5 b 300 m
3.17(1.655 . 10 &5 m2/s)2(300 m) 0.5 . ln c0 ' a b d (9.81 m/s2)(0.267 m)3(20 m)
! 0.24 m3/s
Note that the result from the Swamee–Jain relation is the same (to two significant digits) as that obtained with the Colebrook equation using EES or using our manual iteration technique. Therefore, the simple Swamee–Jain relation can be used with confidence.
8–6
■
MINOR LOSSES
The fluid in a typical piping system passes through various fittings, valves, bends, elbows, tees, inlets, exits, enlargements, and contractions in addition to the pipes. These components interrupt the smooth flow of the fluid and cause additional losses because of the flow separation and mixing they induce. In a typical system with long pipes, these losses are minor compared to the total head loss in the pipes (the major losses) and are called minor losses. Although this is generally true, in some cases the minor losses may be greater than the major losses. This is the case, for example, in systems with several turns and valves in a short distance. The head loss introduced by a completely open valve, for example, may be negligible. But a partially closed valve may cause the largest head loss in the system, as
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348 FLUID MECHANICS Pipe section with valve:
V
1
2 (P1 – P2)valve
Loss coefficient:
Pipe section without valve:
V
1
evidenced by the drop in the flow rate. Flow through valves and fittings is very complex, and a theoretical analysis is generally not plausible. Therefore, minor losses are determined experimentally, usually by the manufacturers of the components. Minor losses are usually expressed in terms of the loss coefficient KL (also called the resistance coefficient), defined as (Fig. 8–32)
2 (P1 – P2)pipe ∆PL = (P1 – P2)valve – (P1 – P2)pipe
FIGURE 8–32 For a constant-diameter section of a pipe with a minor loss component, the loss coefficient of the component (such as the gate valve shown) is determined by measuring the additional pressure loss it causes and dividing it by the dynamic pressure in the pipe.
KL !
hL V /(2g) 2
(8–55)
where hL is the additional irreversible head loss in the piping system caused by insertion of the component, and is defined as hL ! (PL /rg. For example, imagine replacing the valve in Fig. 8–32 with a section of constant diameter pipe from location 1 to location 2. (PL is defined as the pressure drop from 1 to 2 for the case with the valve, (P1 & P2 )valve, minus the pressure drop that would occur in the imaginary straight pipe section from 1 to 2 without the valve, (P1 & P2 )pipe at the same flow rate. While the majority of the irreversible head loss occurs locally near the valve, some of it occurs downstream of the valve due to induced swirling turbulent eddies that are produced in the valve and continue downstream. These eddies “waste” mechanical energy because they are ultimately dissipated into heat while the flow in the downstream section of pipe eventually returns to fully developed conditions. When measuring minor losses in some minor loss components, such as elbows, for example, location 2 must be considerably far downstream (tens of pipe diameters) in order to fully account for the additional irreversible losses due to these decaying eddies. When the pipe diameter downstream of the component changes, determination of the minor loss is even more complicated. In all cases, however, it is based on the additional irreversible loss of mechanical energy that would otherwise not exist if the minor loss component were not there. For simplicity, you may think of the minor loss as occurring locally across the minor loss component, but keep in mind that the component influences the flow for several pipe diameters downstream. By the way, this is the reason why most flow meter manufacturers recommend installing their flow meter at least 10 to 20 pipe diameters downstream of any elbows or valves—this allows the swirling turbulent eddies generated by the elbow or valve to largely disappear and the velocity profile to become fully developed before entering the flow meter. (Most flow meters are calibrated with a fully developed velocity profile at the flow meter inlet, and yield the best accuracy when such conditions also exist in the actual application.) When the inlet diameter equals outlet diameter, the loss coefficient of a component can also be determined by measuring the pressure loss across the component and dividing it by the dynamic pressure, KL ! (PL/(12 rV 2). When the loss coefficient for a component is available, the head loss for that component is determined from Minor loss:
hL ! K L
V2 2g
(8–56)
The loss coefficient, in general, depends on the geometry of the component and the Reynolds number, just like the friction factor. However, it is usually assumed to be independent of the Reynolds number. This is a reasonable approximation since most flows in practice have large Reynolds numbers
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349 CHAPTER 8
and the loss coefficients (including the friction factor) tend to be independent of the Reynolds number at large Reynolds numbers. Minor losses are also expressed in terms of the equivalent length Lequiv, defined as (Fig. 8–33) Equivalent length:
hL ! K L
L equiv V 2 V2 ! f 2g D 2g
→ L equiv !
D KL f
(8–57)
where f is the friction factor and D is the diameter of the pipe that contains the component. The head loss caused by the component is equivalent to the head loss caused by a section of the pipe whose length is Lequiv. Therefore, the contribution of a component to the head loss can be accounted for by simply adding Lequiv to the total pipe length. Both approaches are used in practice, but the use of loss coefficients is more common. Therefore, we will also use that approach in this book. Once all the loss coefficients are available, the total head loss in a piping system is determined from Total head loss (general):
h L, total ! hL, major ' hL, minor
V 2j L i V 2i ! a fi ' a K L, j D i 2g 2g i j
(8–58)
where i represents each pipe section with constant diameter and j represents each component that causes a minor loss. If the entire piping system being analyzed has a constant diameter, Eq. 8–58 reduces to Total head loss (D ! constant):
h L, total ! af
L V2 ' a K Lb D 2g
2 D 1
∆P = P1 – P2 = P3 – P4 3
4 D
FIGURE 8–33 The head loss caused by a component (such as the angle valve shown) is equivalent to the head loss caused by a section of the pipe whose length is the equivalent length.
(8–59)
where V is the average flow velocity through the entire system (note that V ! constant since D ! constant). Representative loss coefficients KL are given in Table 8–4 for inlets, exits, bends, sudden and gradual area changes, and valves. There is considerable uncertainty in these values since the loss coefficients, in general, vary with the pipe diameter, the surface roughness, the Reynolds number, and the details of the design. The loss coefficients of two seemingly identical valves by two different manufacturers, for example, can differ by a factor of 2 or more. Therefore, the particular manufacturer’s data should be consulted in the final design of piping systems rather than relying on the representative values in handbooks. The head loss at the inlet of a pipe is a strong function of geometry. It is almost negligible for well-rounded inlets (KL ! 0.03 for r/D ) 0.2), but increases to about 0.50 for sharp-edged inlets (Fig. 8–34). That is, a sharpedged inlet causes half of the velocity head to be lost as the fluid enters the pipe. This is because the fluid cannot make sharp 90° turns easily, especially at high velocities. As a result, the flow separates at the corners, and the flow is constricted into the vena contracta region formed in the midsection of the pipe (Fig. 8–35). Therefore, a sharp-edged inlet acts like a flow constriction. The velocity increases in the vena contracta region (and the pressure decreases) because of the reduced effective flow area and then decreases as the flow fills the entire cross section of the pipe. There would be negligible loss if the pressure were increased in accordance with Bernoulli’s equation (the velocity head would simply be converted into pressure head). However, this deceleration process is far from ideal and the
Sharp-edged inlet KL = 0.50
Recirculating flow
Well-rounded inlet KL = 0.03
D
r
FIGURE 8–34 The head loss at the inlet of a pipe is almost negligible for well-rounded inlets (KL ! 0.03 for r/D ) 0.2) but increases to about 0.50 for sharp-edged inlets.
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TA B L E 8 – 4 Loss coefficients KL of various pipe components for turbulent flow (for use in the relation hL ! KLV 2/(2g), where V is the average velocity in the pipe that contains the component)* Pipe Inlet Reentrant: KL ! 0.80 (t ** D and I $ 0.1D)
Sharp-edged: KL ! 0.50
Well-rounded (r/D ) 0.2): KL ! 0.03 Slightly rounded (r/D ! 0.1): KL ! 0.12 (see Fig. 8–36)
r
V
D
V
V D
D
t
l
Pipe Exit Reentrant: KL ! a
Sharp-edged: KL ! a
V
Rounded: KL ! a
V
V
Note: The kinetic energy correction factor is a ! 2 for fully developed laminar flow, and a # 1 for fully developed turbulent flow.
Sudden Expansion and Contraction (based on the velocity in the smaller-diameter pipe) Sudden expansion: KL ! a1 &
V
d
d2 2 b D2 0.6
D
0.4
Sudden contraction: See chart.
KL for sudden contraction
KL 0.2
D
d
V
0
0
0.2
0.4
0.6
0.8
1.0
d2/D2
Gradual Expansion and Contraction (based on the velocity in the smaller-diameter pipe) Expansion: Contraction (for u ! 20°): KL ! 0.02 for u ! 20° KL ! 0.30 for d/D ! 0.2 KL ! 0.04 for u ! 45° KL ! 0.25 for d/D ! 0.4 D V d u KL ! 0.07 for u ! 60° KL ! 0.15 for d/D ! 0.6 KL ! 0.10 for d/D ! 0.8
D
u
d
V
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TA B L E 8 – 4 ( C O N C L U D E D ) Bends and Branches 90° smooth bend: Flanged: KL ! 0.3 Threaded: KL ! 0.9
90° miter bend (without vanes): KL ! 1.1
90° miter bend (with vanes): KL ! 0.2
45° threaded elbow: KL ! 0.4
V
V
V
V
180° return bend: Flanged: KL ! 0.2 Threaded: KL ! 1.5
Tee (branch flow): Flanged: KL ! 1.0 Threaded: KL ! 2.0
Tee (line flow): Flanged: KL ! 0.2 Threaded: KL ! 0.9
Threaded union: KL ! 0.08
V
V
45°
V V
Valves Globe valve, fully open: KL ! 10 Angle valve, fully open: KL ! 5 Ball valve, fully open: KL ! 0.05 Swing check valve: KL ! 2
Gate valve, fully open: 1 4 closed: 1 2 closed: 3 4 closed:
KL KL KL KL
! ! ! !
0.2 0.3 2.1 17
* These are representative values for loss coefficients. Actual values strongly depend on the design and manufacture of the components and may differ from the given values considerably (especially for valves). Actual manufacturer’s data should be used in the final design.
Pressure head converted to velocity head
Head
P0 rg
Total head
V 21 2g P1 rg
Pressure head
1
KLV 2/2g
Lost velocity head
V 22 /2g
Remaining velocity head
P2 rg
Remaining pressure head
2
Vena contracta
0
1
Separated flow
2
FIGURE 8–35 Graphical representation of flow contraction and the associated head loss at a sharp-edged pipe inlet.
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352 FLUID MECHANICS 0.5 r 0.4
D
0.3 KL 0.2
FIGURE 8–36 The effect of rounding of a pipe inlet on the loss coefficient. From ASHRAE Handbook of Fundamentals.
Mixing Submerged outlet
Entrained ambient fluid
FIGURE 8–37 All the kinetic energy of the flow is “lost” (turned into thermal energy) through friction as the jet decelerates and mixes with ambient fluid downstream of a submerged outlet.
0.1 0
0
0.05
0.10
0.15
0.20
0.25
r/D
viscous dissipation caused by intense mixing and the turbulent eddies convert part of the kinetic energy into frictional heating, as evidenced by a slight rise in fluid temperature. The end result is a drop in velocity without much pressure recovery, and the inlet loss is a measure of this irreversible pressure drop. Even slight rounding of the edges can result in significant reduction of KL, as shown in Fig. 8–36. The loss coefficient rises sharply (to about KL ! 0.8) when the pipe protrudes into the reservoir since some fluid near the edge in this case is forced to make a 180° turn. The loss coefficient for a submerged pipe exit is often listed in handbooks as KL ! 1. More precisely, however, KL is equal to the kinetic energy correction factor a at the exit of the pipe. Although a is indeed close to 1 for fully developed turbulent pipe flow, it is equal to 2 for fully developed laminar pipe flow. To avoid possible errors when analyzing laminar pipe flow, then, it is best to always set KL ! a at a submerged pipe exit. At any such exit, whether laminar or turbulent, the fluid leaving the pipe loses all of its kinetic energy as it mixes with the reservoir fluid and eventually comes to rest through the irreversible action of viscosity. This is true, regardless of the shape of the exit (Table 8–4 and Fig. 8–37). Therefore, there is no need to round the pipe exits. Piping systems often involve sudden or gradual expansion or contraction sections to accommodate changes in flow rates or properties such as density and velocity. The losses are usually much greater in the case of sudden expansion and contraction (or wide-angle expansion) because of flow separation. By combining the conservation of mass, momentum, and energy equations, the loss coefficient for the case of sudden expansion is approximated as K L ! a1 &
A small 2 b A large
(sudden expansion)
(8–60)
where Asmall and Alarge are the cross-sectional areas of the small and large pipes, respectively. Note that KL ! 0 when there is no area change (Asmall ! Alarge) and KL ! 1 when a pipe discharges into a reservoir (Alarge )) Asmall). No such relation exists for a sudden contraction, and the KL values in that case can be read from the chart in Table 8–4. The losses due to expansion and contraction can be reduced significantly by installing conical gradual area changers (nozzles and diffusers) between the small and large
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353 CHAPTER 8
pipes. The KL values for representative cases of gradual expansion and contraction are given in Table 8–4. Note that in head loss calculations, the velocity in the small pipe is to be used as the reference velocity in Eq. 8–56. Losses during expansion are usually much higher than the losses during contraction because of flow separation. Piping systems also involve changes in direction without a change in diameter, and such flow sections are called bends or elbows. The losses in these devices are due to flow separation (just like a car being thrown off the road when it enters a turn too fast) on the inner side and the swirling secondary flows caused by different path lengths. The losses during changes of direction can be minimized by making the turn “easy” on the fluid by using circular arcs (like the 90° elbow) instead of sharp turns (like miter bends) (Fig. 8–38). But the use of sharp turns (and thus suffering a penalty in loss coefficient) may be necessary when the turning space is limited. In such cases, the losses can be minimized by utilizing properly placed guide vanes to help the flow turn in an orderly manner without being thrown off the course. The loss coefficients for some elbows and miter bends as well as tees are given in Table 8–4. These coefficients do not include the frictional losses along the pipe bend. Such losses should be calculated as in straight pipes (using the length of the centerline as the pipe length) and added to other losses. Valves are commonly used in piping systems to control the flow rates by simply altering the head loss until the desired flow rate is achieved. For valves it is desirable to have a very low loss coefficient when they are fully open so that they cause minimal head loss during full-load operation. Several different valve designs, each with its own advantages and disadvantages, are in common use today. The gate valve slides up and down like a gate, the globe valve closes a hole placed in the valve, the angle valve is a globe valve with a 90° turn, and the check valve allows the fluid to flow only in one direction like a diode in an electric circuit. Table 8–4 lists the representative loss coefficients of the popular designs. Note that the loss coefficient increases drastically as a valve is closed (Fig. 8–39). Also, the deviation in the loss coefficients for different manufacturers is greatest for valves because of their complex geometries.
EXAMPLE 8–6
Flanged elbow KL = 0.3
Sharp turn KL = 1.1
FIGURE 8–38 The losses during changes of direction can be minimized by making the turn “easy” on the fluid by using circular arcs instead of sharp turns.
A globe valve
V1
V2 V2 = V1 Vconstriction > V1
Constriction
FIGURE 8–39 The large head loss in a partially closed valve is due to irreversible deceleration, flow separation, and mixing of high-velocity fluid coming from the narrow valve passage.
Head Loss and Pressure Rise during Gradual Expansion
A 6-cm-diameter horizontal water pipe expands gradually to a 9-cm-diameter pipe (Fig. 8–40). The walls of the expansion section are angled 30° from the horizontal. The average velocity and pressure of water before the expansion section are 7 m/s and 150 kPa, respectively. Determine the head loss in the expansion section and the pressure in the larger-diameter pipe.
SOLUTION A horizontal water pipe expands gradually into a larger-diameter pipe. The head loss and pressure after the expansion are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The flow at sections 1 and 2 is fully developed and turbulent with a1 ! a2 % 1.06. Properties We take the density of water to be r ! 1000 kg/m3. The loss coefficient for gradual expansion of u ! 60° total included angle is KL ! 0.07.
1
6 cm
9 cm
2
Water 7 m/s 150 kPa
FIGURE 8–40 Schematic for Example 8–6.
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354 FLUID MECHANICS
Analysis Noting that the density of water remains constant, the downstream velocity of water is determined from conservation of mass to be
# # m1 ! m2 →
rV1 A 1 ! rV2 A 2 → V2 !
V2 !
A1 D 21 V1 ! 2 V1 A2 D2
(0.06 m)2 (7 m/s) ! 3.11 m/s (0.09 m)2
Then the irreversible head loss in the expansion section becomes
hL ! K L
V 21 (7 m/s)2 ! (0.07) ! 0.175 m 2g 2(9.81 m/s2)
Noting that z1 ! z2 and there are no pumps or turbines involved, the energy equation for the expansion section can be expressed in terms of heads as
0 V 21 P1 S ' a1 ' z 1 ' hpump, u rg 2g
!
0 P2 V 22 ¡ e ' hL ' a2 ' z 2 ' hturbine, 2g rg
→
P1 V 21 P2 V 22 ' a1 ' a2 ! ' hL rg rg 2g 2g
Solving for P2 and substituting,
a 1V 21 & a 2V 22 P2 ! P1 ' r e & ghLf ! (150 kPa) ' (1000 kg/m3) 2 1.06(7 m/s)2 & 1.06(3.11 m/s)2 .e & (9.81 m/s2)(0.175 m)f 2 1 kN 1 kPa .a b ba 1000 kg , m/s 1 kN/m2
! 169 kPa
Therefore, despite the head (and pressure) loss, the pressure increases from 150 to 169 kPa after the expansion. This is due to the conversion of dynamic pressure to static pressure when the average flow velocity is decreased in the larger pipe. Discussion It is common knowledge that higher pressure upstream is necessary to cause flow, and it may come as a surprise to you that the downstream pressure has increased after the expansion, despite the loss. This is because the flow is driven by the sum of the three heads that comprise the total head (namely, the pressure head, velocity head, and elevation head). During flow expansion, the higher velocity head upstream is converted to pressure head downstream, and this increase outweighs the nonrecoverable head loss. Also, you may be tempted to solve this problem using the Bernoulli equation. Such a solution would ignore the head (and the associated pressure) loss and result in an incorrect higher pressure for the fluid downstream.
8–7 FIGURE 8–41 A piping network in an industrial facility. Courtesy UMDE Engineering, Contracting, and Trading. Used by permission.
■
PIPING NETWORKS AND PUMP SELECTION
Most piping systems encountered in practice such as the water distribution systems in cities or commercial or residential establishments involve numerous parallel and series connections as well as several sources (supply of fluid into the system) and loads (discharges of fluid from the system) (Fig. 8–41). A piping project may involve the design of a new system or the
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355 CHAPTER 8
expansion of an existing system. The engineering objective in such projects is to design a piping system that will deliver the specified flow rates at specified pressures reliably at minimum total (initial plus operating and maintenance) cost. Once the layout of the system is prepared, the determination of the pipe diameters and the pressures throughout the system, while remaining within the budget constraints, typically requires solving the system repeatedly until the optimal solution is reached. Computer modeling and analysis of such systems make this tedious task a simple chore. Piping systems typically involve several pipes connected to each other in series and/or in parallel, as shown in Figs. 8–42 and 8–43. When the pipes are connected in series, the flow rate through the entire system remains constant regardless of the diameters of the individual pipes in the system. This is a natural consequence of the conservation of mass principle for steady incompressible flow. The total head loss in this case is equal to the sum of the head losses in individual pipes in the system, including the minor losses. The expansion or contraction losses at connections are considered to belong to the smaller-diameter pipe since the expansion and contraction loss coefficients are defined on the basis of the average velocity in the smaller-diameter pipe. For a pipe that branches out into two (or more) parallel pipes and then rejoins at a junction downstream, the total flow rate is the sum of the flow rates in the individual pipes. The pressure drop (or head loss) in each individual pipe connected in parallel must be the same since (P ! PA & PB and the junction pressures PA and PB are the same for all the individual pipes. For a system of two parallel pipes 1 and 2 between junctions A and B with negligible minor losses, this can be expressed as →
h L, 1 ! hL, 2
f1
A
B
1
2 fA, LA, DA
fB, LB, DB
⋅ ⋅ VA = VB hL, 1-2 = hL, A + hL, B
FIGURE 8–42 For pipes in series, the flow rate is the same in each pipe, and the total head loss is the sum of the head losses in individual pipes.
L 1 V 21 L 2 V 22 ! f2 D 1 2g D 2 2g
Then the ratio of the average velocities and the flow rates in the two parallel pipes become f 2 L 2 D 1 1/2 V1 !a b V2 f1 L 1 D2
# A c, 1V1 D 21 f 2 L 2 D 1 1/2 V1 ! 2a b # ! A c, 2V2 D2 f1 L 1 D2 V2
and
Therefore, the relative flow rates in parallel pipes are established from the requirement that the head loss in each pipe be the same. This result can be extended to any number of pipes connected in parallel. The result is also valid for pipes for which the minor losses are significant if the equivalent lengths for components that contribute to minor losses are added to the pipe f1, L1, D1 PA
A
PB < PA
A
B
B
f2, L2, D2 hL, 1 = hL, 2 ⋅ ⋅ ⋅ ⋅ VA = V1 + V2 = VB
FIGURE 8–43 For pipes in parallel, the head loss is the same in each pipe, and the total flow rate is the sum of the flow rates in individual pipes.
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356 FLUID MECHANICS
length. Note that the flow rate in one of the parallel branches is proportional to its diameter to the power 5/2 and is inversely proportional to the square root of its length and friction factor. The analysis of piping networks, no matter how complex they are, is based on two simple principles: 1. Conservation of mass throughout the system must be satisfied. This is done by requiring the total flow into a junction to be equal to the total flow out of the junction for all junctions in the system. Also, the flow rate must remain constant in pipes connected in series regardless of the changes in diameters. 2. Pressure drop (and thus head loss) between two junctions must be the same for all paths between the two junctions. This is because pressure is a point function and it cannot have two values at a specified point. In practice this rule is used by requiring that the algebraic sum of head losses in a loop (for all loops) be equal to zero. (A head loss is taken to be positive for flow in the clockwise direction and negative for flow in the counterclockwise direction.) Therefore, the analysis of piping networks is very similar to the analysis of electric circuits, with flow rate corresponding to electric current and pressure corresponding to electric potential. However, the situation is much more complex here since, unlike the electrical resistance, the “flow resistance” is a highly nonlinear function. Therefore, the analysis of piping networks requires the simultaneous solution of a system of nonlinear equations. The analysis of such systems is beyond the scope of this introductory text.
Piping Systems with Pumps and Turbines
When a piping system involves a pump and/or turbine, the steady-flow energy equation on a unit-mass basis can be expressed as (see Section 5–7) P1 V 21 P2 V 22 ' a1 ' a2 ' gz 1 ' wpump, u ! ' gz 2 ' wturbine, e ' ghL (8–61) r r 2 2
It can also be expressed in terms of heads as P1 V 21 P2 V 22 ' a1 ' a2 ' z 1 ' hpump, u ! ' z 2 ' hturbine, e ' hL rg rg 2g 2g
(8–62)
where hpump, u ! wpump, u /g is the useful pump head delivered to the fluid, hturbine, e ! wturbine, e /g is the turbine head extracted from the fluid, a is the kinetic energy correction factor whose value is nearly 1 for most (turbulent) flows encountered in practice, and hL is the total head loss in piping (including the minor losses if they are significant) between points 1 and 2. The pump head is zero if the piping system does not involve a pump or a fan, the turbine head is zero if the system does not involve a turbine, and both are zero if the system does not involve any mechanical work-producing or work-consuming devices. Many practical piping systems involve a pump to move a fluid from one reservoir to another. Taking points 1 and 2 to be at the free surfaces of the reservoirs, the energy equation in this case reduces for the useful pump head required to (Fig. 8–44) h pump, u ! (z 2 & z 1) ' hL
(8–63)
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357 CHAPTER 8 2
1
z1
Control volume boundary
z2
Pump hpump, u = (z2 – z1) + hL ⋅ ⋅ Wpump, u = rVghpump, u
FIGURE 8–44 When a pump moves a fluid from one reservoir to another, the useful pump head requirement is equal to the elevation difference between the two reservoirs plus the head loss.
since the velocities at free surfaces are negligible and the pressures are at atmospheric pressure. Therefore, the useful pump head is equal to the elevation difference between the two reservoirs plus the head loss. If the head loss is negligible compared to z2 & z1, the useful pump head is simply equal to the elevation difference between the two reservoirs. In the case of z1 ) z2 (the first reservoir being at a higher elevation than the second one) with no pump, the flow is driven by gravity at a flow rate that causes a head loss equal to the elevation difference. A similar argument can be given for the turbine head for a hydroelectric power plant by replacing hpump, u in Eq. 8–63 by &hturbine, e. Once the useful pump head is known, the mechanical power that needs to be delivered by the pump to the fluid and the electric power consumed by the motor of the pump for # a specified flow rate are determined from # rV ghpump, u # Wpump, shaft ! h pump
and
rV ghpump, u # Welect ! h pump–motor
Motor hmotor = 0.90
(8–64)
where hpump–motor is the efficiency of the pump–motor combination, which is the product of the pump and the motor efficiencies (Fig. 8–45). The pump–motor efficiency is defined as the ratio of the net mechanical energy delivered to the fluid by the pump to the electric energy consumed by the motor of the pump, and it usually ranges between 50 and 85 percent. The head loss of a piping system increases (usually quadratically) with the flow rate. A plot of required useful pump head hpump, u as a function of flow rate is called the system (or demand) curve. The head produced by a pump is not a constant either. Both the pump head and the pump efficiency vary with the flow rate, and pump manufacturers supply this variation in tabular or graphical form, as shown. in Fig. 8–46. These experimentally determined hpump, u and hpump, u versus V curves are called characteristic (or supply or performance) curves. Note that the flow rate of a pump increases as the required head decreases. The intersection point of the pump head curve with the vertical axis typically represents the maximum head the pump can provide, while the intersection point with the horizontal axis indicates the maximum flow rate (called the free delivery) that the pump can supply. The efficiency of a pump is sufficiently high for a certain range of head and flow rate combination. Therefore, a pump that can supply the required head and flow rate is not necessarily a good choice for a piping system unless the efficiency of the pump at those conditions is sufficiently high. The pump installed in a piping system will operate at the point where the system curve and the characteristic curve intersect. This point of intersection is called the operating point, as shown in Fig. 8–46. The useful head
Liquid out Liquid in
Pump hpump = 0.70
hpump–motor = hpumphmotor = 0.70 . 0.90 = 0.63
FIGURE 8–45 The efficiency of the pump–motor combination is the product of the pump and the motor efficiencies. Courtesy Yunus Çengel
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358 FLUID MECHANICS Pump exit is closed to produce maximum head
hpump
Head, m
FIGURE 8–46 Characteristic pump curves for centrifugal pumps, the system curve for a piping system, and the operating point.
80
30
60 Operating point
20
10 System curve 0
0
1
2
Supply curve
4
3
40
20
5
6
0
Pump efficiency, % hpump
hpump, u 40
100
No pipe is attached to the pump (no load to maximize flow rate)
Flow rate, m3/s
produced by the pump at this point matches the head requirements of the system at that flow rate. Also, the efficiency of the pump during operation is the value corresponding to that flow rate. EXAMPLE 8–7
Pumping Water through Two Parallel Pipes
Water at 20°C is to be pumped from a reservoir (zA ! 5 m) to another reservoir at a higher elevation (zB ! 13 m) through two 36-m-long pipes connected in parallel, as shown in Fig. 8–47. The pipes are made of commercial steel, and the diameters of the two pipes are 4 and 8 cm. Water is to be pumped by a 70 percent efficient motor–pump combination that draws 8 kW of electric power during operation. The minor losses and the head loss in pipes that connect the parallel pipes to the two reservoirs are considered to be negligible. Determine the total flow rate between the reservoirs and the flow rate through each of the parallel pipes.
SOLUTION The pumping power input to a piping system with two parallel pipes is given. The flow rates are to be determined.
L1 = 36 m D1 = 4 cm
A
Control volume boundary
zA = 5 m 1
FIGURE 8–47 The piping system discussed in Example 8–7.
Pump
2
D2 = 8 cm L2 = 36 m
B
zB = 13 m
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359 CHAPTER 8
Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The elevations of the reservoirs remain constant. 4 The minor losses and the head loss in pipes other than the parallel pipes are said to be negligible. 5 Flows through both pipes are turbulent (to be verified). Properties The density and dynamic viscosity of water at 20°C are r ! 998 kg/m3 and m ! 1.002 . 10&3 kg/m · s. The roughness of commercial steel pipe is e ! 0.000045 m. Analysis This problem cannot be solved directly since the velocities (or flow rates) in the pipes are not known. Therefore, we would normally use a trialand-error approach here. However, nowadays equation solvers such as EES are widely available, and thus we will simply set up the equations to be solved by an equation solver. The useful head supplied by the pump to the fluid is determined from
# rV ghpump, u # W elect ! h pump&motor
→
8000 W !
# (998 kg/m3)V (9.81 m/s2)hpump, u 0.70
(1)
We choose points A and B at the free surfaces of the two reservoirs. Noting that the fluid at both points is open to the atmosphere (and thus PA ! PB ! Patm) and that the fluid velocities at both points are zero (VA ! VB ! 0), the energy equation for a control volume between these two points simplifies to
0 0 V 2A Q PB V 2B Q PA ' aA ' aB ' z A ' hpump, u ! ' z B ' hL → hpump, u rg rg 2g 2g ! (z B & z A) ' hL or
hpump, u ! (13 & 5) ' hL
(2)
where
hL ! hL, 1 ! hL, 2
(3)(4)
We designate the 4-cm-diameter pipe by 1 and the 8-cm-diameter pipe by 2. The average velocity, the Reynolds number, the friction factor, and the head loss in each pipe are expressed as
# # V1 V1 ! V1 ! A c, 1 pD 21/4 # # V2 V2 V2 ! ! A c, 2 pD 22/4
→ →
Re1 !
rV1D 1 m
→
Re1 !
Re2 !
rV2D 2 m
→
Re2 !
# V1 V1 ! p(0.04 m)2/4 # V2 V2 ! p(0.08 m)2/4
(998 kg/m3)V1(0.04 m) 1.002 . 10 &3 kg/m , s (998 kg/m3)V2(0.08 m) 1.002 . 10 &3 kg/m , s
(5)
(6)
(7)
(8)
1
e/D 1 2.51 ' ! &2.0 log a b 3.7 2f 1 Re1 2f 1 →
1
0.000045 2.51 ! &2.0 loga ' b 3.7 . 0.04 2f 1 Re1 2f 1
(9)
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360 FLUID MECHANICS
1 2f 2
e/D 2 2.51 b ' 3.7 Re2 2f 2
! &2.0 loga 1
→ h L, 1 ! f 1
0.000045 2.51 ! &2.0 loga ' b 3.7 . 0.08 2f 2 Re2 2f 2
L 1 V 21 D 1 2g
L 2 V 22 D 2 2g # # # V ! V1 ' V2
h L, 2 ! f 2
(10)
→
hL, 1 ! f 1
V 21 36 m 0.04 m 2(9.81 m/s2)
(11)
→
hL, 2 ! f 2
V 22 36 m 0.08 m 2(9.81 m/s2)
(12) (13)
This is a system of 13 equations in 13 unknowns, and their simultaneous solution by an equation solver gives
# V ! 0.0300 m3/s,
# V 1 ! 0.00415 m3/s,
# V 2 ! 0.0259 m3/s
V1 ! 3.30 m/s, V2 ! 5.15 m/s, h L ! hL, 1 ! hL, 2 ! 11.1 m, hpump ! 19.1 m Re1 ! 131,600,
Re2 ! 410,000,
f 1 ! 0.0221,
f 2 ! 0.0182
Note that Re ) 4000 for both pipes, and thus the assumption of turbulent flow is verified. Discussion The two parallel pipes are identical, except the diameter of the first pipe is half the diameter of the second one. But only 14 percent of the water flows through the first pipe. This shows the strong dependence of the flow rate (and the head loss) on diameter. Also, it can be shown that if the free surfaces of the two reservoirs were at the same elevation (and thus zA ! zB), the flow rate would increase by 20 percent from 0.0300 to 0.0361 m3/s. Alternately, if the reservoirs were as given but the irreversible head losses were negligible, the flow rate would become 0.0715 m3/s (an increase of 138 percent).
EXAMPLE 8–8
Gravity-Driven Water Flow in a Pipe
Water at 10°C flows from a large reservoir to a smaller one through a 5-cmdiameter cast iron piping system, as shown in Fig. 8–48. Determine the elevation z1 for a flow rate of 6 L/s.
z1 = ?
1
Sharp-edged entrance, KL = 0.5 Standard elbow, flanged, KL = 0.3 D = 5 cm 2
FIGURE 8–48 The piping system discussed in Example 8–8.
z2 = 4 m
Gate valve, fully open KL = 0.2
9m Control volume boundary 80 m
Exit, KL = 1.06
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361 CHAPTER 8
SOLUTION The flow rate through a piping system connecting two reservoirs is given. The elevation of the source is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The elevations of the reservoirs remain constant. 3 There are no pumps or turbines in the line. Properties The density and dynamic viscosity of water at 10°C are r ! 999.7 kg/m3 and m ! 1.307 . 10&3 kg/m · s. The roughness of cast iron pipe is e ! 0.00026 m. Analysis The piping system involves 89 m of piping, a sharp-edged entrance (KL ! 0.5), two standard flanged elbows (KL ! 0.3 each), a fully open gate valve (KL ! 0.2), and a submerged exit (KL ! 1.06). We choose points 1 and 2 at the free surfaces of the two reservoirs. Noting that the fluid at both points is open to the atmosphere (and thus P1 ! P2 ! Patm) and that the fluid velocities at both points are zero (V1 ! V2 ! 0), the energy equation for a control volume between these two points simplifies to 0 0 V 21 Q P2 V 22 Q P1 ' a1 ' a2 ' z1 ! ' z 2 ' hL rg rg 2g 2g where
hL ! hL, total ! hL, major ' hL, minor ! af
→
z 1 ! z 2 ' hL
L V2 ' a K Lb D 2g
since the diameter of the piping system is constant. The average velocity in the pipe and the Reynolds number are
# # V V 0.006 m3/s V ! ! ! ! 3.06 m/s A c pD2/4 p(0.05 m)2/4
Re !
rVD (999.7 kg/m3)(3.06 m/s)(0.05 m) ! ! 117,000 m 1.307 . 10 &3 kg/m , s
The flow is turbulent since Re ) 4000. Noting that e/D ! 0.00026/0.05 ! 0.0052, the friction factor can be determined from the Colebrook equation (or the Moody chart), 1
! &2.0 loga
e/D 2.51 ' b 3.7 Re 2f
→
1
0.0052 2.51 ' b 3.7 117,0002f
! &2.0 loga
2f 2f It gives f ! 0.0315. The sum of the loss coefficients is
a K L ! K L, entrance ' 2K L, elbow ' K L, valve ' K L, exit ! 0.5 ' 2 . 0.3 ' 0.2 ' 1.06 ! 2.36
Then the total head loss and the elevation of the source become
hL ! af
(3.06 m/s)2 L V2 89 m ' a K Lb ! a0.0315 ' 2.36b ! 27.9 m D 2g 0.05 m 2(9.81 m/s2)
z 1 ! z 2 ' hL ! 4 ' 27.9 ! 31.9 m
Therefore, the free surface of the first reservoir must be 31.9 m above the ground level to ensure water flow between the two reservoirs at the specified rate. Discussion Note that fL/D ! 56.1 in this case, which is about 24 times the total minor loss coefficient. Therefore, ignoring the sources of minor losses in this case would result in about 4 percent error. It can be shown that the total head loss would be 35.9 m (instead of 27.9 m) if the valve were three-fourths closed, and it would drop to 24.8 m if the pipe between the two reservoirs were straight at the ground level (thus
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362 FLUID MECHANICS
eliminating the elbows and the vertical section of the pipe). The head loss could be reduced further (from 24.8 to 24.6 m) by rounding the entrance. The head loss can be reduced significantly (from 27.9 to 16.0 m) by replacing the cast iron pipes by smooth pipes such as those made of plastic.
EXAMPLE 8–9
Effect of Flushing on Flow Rate from a Shower
The bathroom plumbing of a building consists of 1.5-cm-diameter copper pipes with threaded connectors, as shown in Fig. 8–49. (a) If the gage pressure at the inlet of the system is 200 kPa during a shower and the toilet reservoir is full (no flow in that branch), determine the flow rate of water through the shower head. (b) Determine the effect of flushing of the toilet on the flow rate through the shower head. Take the loss coefficients of the shower head and the reservoir to be 12 and 14, respectively.
SOLUTION The cold-water plumbing system of a bathroom is given. The flow rate through the shower and the effect of flushing the toilet on the flow rate are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The flow is turbulent and fully developed. 3 The reservoir is open to the atmosphere. 4 The velocity heads are negligible. Properties The properties of water at 20°C are r ! 998 kg/m3, m ! 1.002 . 10&3 kg/m · s, and n ! m/r ! 1.004 . 10&6 m2/s. The roughness of copper pipes is e ! 1.5 . 10&6 m. Analysis This is a problem of the second type since it involves the determination of the flow rate for a specified pipe diameter and pressure drop. The solution involves an iterative approach since the flow rate (and thus the flow velocity) is not known. (a) The piping system of the shower alone involves 11 m of piping, a tee with line flow (KL ! 0.9), two standard elbows (KL ! 0.9 each), a fully open globe valve (KL ! 10), and a shower head (KL ! 12). Therefore, &KL ! 0.9 ' 2 . 0.9 ' 10 ' 12 ! 24.7. Noting that the shower head is open to the atmosphere, and the velocity heads are negligible, the energy equation for a control volume between points 1 and 2 simplifies to V 21 P2 V 22 P1 ' a1 ' a2 ' z 1 ' hpump, u ! ' z 2 ' hturbine, e ' hL 2g 2g rg rg →
P1, gage rg
! (z 2 & z 1) ' hL
Shower head KL = 12 Toilet reservoir with float KL = 14
2
3 2m 1m KL = 2
Cold water
FIGURE 8–49 Schematic for Example 8–9.
1
5m
KL = 0.9
KL = 10
4m
Globe valve, fully open KL = 10
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363 CHAPTER 8
Therefore, the head loss is
hL !
200,000 N/m2 & 2 m ! 18.4 m (998 kg/m3)(9.81 m/s2)
Also,
hL ! af
L V2 ' a K Lb D 2g
→
18.4 ! af
11 m V2 ' 24.7b 0.015 m 2(9.81 m/s2)
since the diameter of the piping system is constant. The average velocity in the pipe, the Reynolds number, and the friction factor are
# # V V ! V ! A c pD2/4
Re !
VD n
→
# V V! p(0.015 m)2/4
→
V(0.015 m) 1.004 . 10 &6 m2/s
Re !
1
2.51 e/D ' ! &2.0 log a b 3.7 Re2f 2f 1
→
2f
1.5 . 10 &6 m 2.51 b ' 3.7(0.015 m) Re 2f
! &2.0 loga
This is a set of four equations with four unknowns, and solving them with an equation solver such as EES gives
# V ! 0.00053 m3/s,
f ! 0.0218, V ! 2.98 m/s,
and
Re ! 44,550
Therefore, the flow rate of water through the shower head is 0.53 L/s. (b) When the toilet is flushed, the float moves and opens the valve. The discharged water starts to refill the reservoir, resulting in parallel flow after the tee connection. The head loss and minor loss coefficients for the shower branch were determined in (a) to be hL, 2 ! 18.4 m and 3KL, 2 ! 24.7, respectively. The corresponding quantities for the reservoir branch can be determined similarly to be
hL, 3 !
200,000 N/m2 & 1 m ! 19.4 m (998 kg/m3)(9.81 m/s2)
3K L, 3 ! 2 ' 10 ' 0.9 ' 14 ! 26.9 The relevant equations in this case are
# # # V1 ! V2 ' V3
hL, 2 ! f 1
V 22 V 21 5m 6m ' 24.7b ' af 2 ! 18.4 2 0.015 m 2(9.81 m/s ) 0.015 m 2(9.81 m/s2)
V 23 V 21 5m 1m ' 26.9b ' af 3 ! 19.4 2 0.015 m 2(9.81 m/s ) 0.015 m 2(9.81 m/s2) # # # V3 V1 V2 V1 ! , V2 ! , V3 ! p(0.015 m)2/4 p(0.015 m)2/4 p(0.015 m)2/4
h L, 3 ! f 1
Re1 !
V3(0.015 m) V1(0.015 m) V2(0.015 m) , Re2 ! , Re3 ! &6 2 &6 2 1.004 . 10 m /s 1.004 . 10 m /s 1.004 . 10 &6m2/s
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364 FLUID MECHANICS
1 2f 1 1 2f 2 1 2f 3
1.5 . 10 &6 m 2.51 b ' 3.7(0.015 m) Re1 2f 1
! &2.0 loga
1.5 . 10 &6 m 2.51 ' b 3.7(0.015 m) Re2 2f 2
! &2.0 loga
1.5 . 10 &6 m 2.51 b ' 3.7(0.015 m) Re3 2f 3
! &2.0 loga
Solving these 12 equations in 12 unknowns simultaneously using an equation solver, the flow rates are determined to be
# V 1 ! 0.00090 m3/s,
FIGURE 8–50 Flow rate of cold water through a shower may be affected significantly by the flushing of a nearby toilet.
# V 2 ! 0.00042 m3/s,
# and V 3 ! 0.00048 m3/s
Therefore, the flushing of the toilet reduces the flow rate of cold water through the shower by 21 percent from 0.53 to 0.42 L/s, causing the shower water to suddenly get very hot (Fig. 8–50). Discussion If the velocity heads were considered, the flow rate through the shower would be 0.43 instead of 0.42 L/s. Therefore, the assumption of negligible velocity heads is reasonable in this case. Note that a leak in a piping system will cause the same effect, and thus an unexplained drop in flow rate at an end point may signal a leak in the system.
8–8
■
FLOW RATE AND VELOCITY MEASUREMENT
A major application area of fluid mechanics is the determination of the flow rate of fluids, and numerous devices have been developed over the years for the purpose of flow metering. Flowmeters range widely in their level of sophistication, size, cost, accuracy, versatility, capacity, pressure drop, and the operating principle. We give an overview of the meters commonly used to measure the flow rate of liquids and gases flowing through pipes or ducts. We limit our consideration to incompressible flow. Some flowmeters measure the flow rate directly by discharging and recharging a measuring chamber of known volume continuously and keeping track of the number of discharges per unit time. But most flowmeters measure the flow rate indirectly—they measure the average velocity V or a quantity that is related to average . velocity such as pressure and drag, and determine the volume flow rate V from # V ! VA c
(8–65)
where Ac is the cross-sectional area of flow. Therefore, measuring the flow rate is usually done by measuring flow velocity, and most flowmeters are simply velocimeters used for the purpose of metering flow. The velocity in a pipe varies from zero at the wall to a maximum at the center, and it is important to keep this in mind when taking velocity measurements. For laminar flow, for example, the average velocity is half the centerline velocity. But this is not the case in turbulent flow, and it may be necessary to take the weighted average of several local velocity measurements to determine the average velocity.
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365 CHAPTER 8
The flow rate measurement techniques range from very crude to very elegant. The flow rate of water through a garden hose, for example, can be measured simply by collecting the water in a bucket of known volume and dividing the amount collected by the collection time (Fig. 8–51). A crude way of estimating the flow velocity of a river is to drop a float on the river and measure the drift time between two specified locations. At the other extreme, some flowmeters use the propagation of sound in flowing fluids while others use the electromotive force generated when a fluid passes through a magnetic field. In this section we discuss devices that are commonly used to measure velocity and flow rate, starting with the Pitot-static probe introduced in Chap. 5.
Pitot and Pitot-Static Probes
Pitot probes (also called Pitot tubes) and Pitot-static probes, named after the French engineer Henri de Pitot (1695–1771), are widely used for flow rate measurement. A Pitot probe is just a tube with a pressure tap at the stagnation point that measures stagnation pressure, while a Pitot-static probe has both a stagnation pressure tap and several circumferential static pressure taps and it measures both stagnation and static pressures (Figs. 8–52 and 8–53). Pitot was the first person to measure velocity with the upstream pointed tube, while French engineer Henry Darcy (1803–1858) developed most of the features of the instruments we use today, including the use of small openings and the placement of the static tube on the same assembly. Therefore, it is more appropriate to call the Pitot-static probes Pitot–Darcy probes. The Pitot-static probe measures local velocity by measuring the pressure difference in conjunction with the Bernoulli equation. It consists of a slender double-tube aligned with the flow and connected to a differential pressure meter. The inner tube is fully open to flow at the nose, and thus it measures the stagnation pressure at that location (point 1). The outer tube is sealed at the nose, but it has holes on the side of the outer wall (point 2) and thus it measures the static pressure. For incompressible flow with sufficiently high velocities (so that the frictional effects between points 1 and 2 are negligible), the Bernoulli equation is applicable and can be expressed as P1 V 21 P2 V 22 ' ' ' z1 ! ' z2 rg 2g rg 2g
V
Pitot probe
Stagnation pressure
Garden hose
Bucket
FIGURE 8–51 A primitive (but fairly accurate) way of measuring the flow rate of water through a garden hose involves collecting water in a bucket and recording the collection time.
(8–66)
Static pressure
To static pressure meter
(a)
Nozzle
Pitot-static probe
V Stagnation pressure
To stagnation pressure meter
Stopwatch
To stagnation pressure meter (b)
FIGURE 8–52 (a) A Pitot probe measures stagnation pressure at the nose of the probe, while (b) a Pitot-static probe measures both stagnation pressure and static pressure, from which the flow speed can be calculated.
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366 FLUID MECHANICS
Noting that z1 ≅ z2 since the static pressure holes of the Pitot-static probe are arranged circumferentially around the tube and V1 ! 0 because of the stagnation conditions, the flow velocity V ! V2 becomes
Pitot-static probe
Flow Stagnation pressure, P1
Wind tunnel wall
Flexible tubing P1 – P2 Differential pressure transducer or manometer to measure P1 – P2
FIGURE 8–53 Measuring flow velocity with a Pitotstatic probe. (A manometer may also be used in place of the differential pressure transducer.)
FIGURE 8–54 Close-up of a Pitot-static probe, showing the stagnation pressure hole and two of the five static circumferential pressure holes. Photo by Po-Ya Abel Chuang.
Obstruction
d
2
D
FIGURE 8–55 Flow through a constriction in a pipe.
B
2(P1 & P2) r
(8–67)
which is known as the Pitot formula. If the velocity is measured at a location where the local velocity is equal to. the average flow velocity, the volume flow rate can be determined from V ! VAc. The Pitot-static probe is a simple, inexpensive, and highly reliable device since it has no moving parts (Fig. 8–54). It also causes very small pressure drop and usually does not disturb the flow appreciably. However, it is important that it be properly aligned with the flow to avoid significant errors that may be caused by misalignment. Also, the difference between the static and stagnation pressures (which is the dynamic pressure) is proportional to the density of the fluid and the square of the flow velocity. It can be used to measure velocity in both liquids and gases. Noting that gases have low densities, the flow velocity should be sufficiently high when the Pitot-static probe is used for gas flow such that a measurable dynamic pressure develops.
Obstruction Flowmeters: Orifice, Venturi, and Nozzle Meters
Consider incompressible steady flow of a fluid in a horizontal pipe of diameter D that is constricted to a flow area of diameter d, as shown in Fig. 8–55. The mass balance and the Bernoulli equations between a location before the constriction (point 1) and the location where constriction occurs (point 2) can be written as Mass balance:
1
V!
Pitot formula:
Static pressure, P2
# V ! A 1V1 ! A 2V2 →
Bernoulli equation (z1 ! z2):
V1 ! (A 2/A 1)V2 ! (d/D)2V2
P1 V 21 P2 V 22 ' ' ! rg 2g rg 2g
(8–68)
(8–69)
Combining Eqs. 8–68 and 8–69 and solving for velocity V2 gives Obstruction (with no loss):
V2 !
2(P1 & P2) B r(1 & b4)
(8–70)
where b ! d/D is .the diameter ratio. Once V2 is known, the flow rate can be determined from V ! A2V2 ! (pd 2/4)V2. This simple analysis shows that the flow rate through a pipe can be determined by constricting the flow and measuring the decrease in pressure due to the increase in velocity at the constriction site. Noting that the pressure drop between two points along the flow can be measured easily by a differential pressure transducer or manometer, it appears that a simple flow rate measurement device can be built by obstructing the flow. Flowmeters based on this principle are called obstruction flowmeters and are widely used to measure flow rates of gases and liquids. The velocity in Eq. 8–70 is obtained by assuming no loss, and thus it is the maximum velocity that can occur at the constriction site. In reality, some pressure losses due to frictional effects are inevitable, and thus the velocity will be less. Also, the fluid stream will continue to contract past the
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367 CHAPTER 8
obstruction, and the vena contracta area is less than the flow area of the obstruction. Both losses can be accounted for by incorporating a correction factor called the discharge coefficient Cd whose value (which is less than 1) is determined experimentally. Then the flow rate for obstruction flowmeters can be expressed as Obstruction flowmeters:
# V ! A 0C d
2(P1 & P2) B r(1 & b 4)
D
d
(8–71)
(a) Orifice meter
where A0 ! A2 ! pd 2/4 is the cross-sectional area of the hole and b ! d/D is the ratio of hole diameter to pipe diameter. The value of Cd depends on both b and the Reynolds number Re ! V1D/n, and charts and curve-fit correlations for Cd are available for various types of obstruction meters. Of the numerous types of obstruction meters available, those most widely used are orifice meters, flow nozzles, and Venturi meters (Fig. 8–56). The experimentally determined data for discharge coefficients are expressed as (Miller, 1997)
D
Orifice meters:
Cd ! 0.5959 ' 0.0312b2.1 & 0.184b8 '
Nozzle meters:
Cd ! 0.9975 &
91.71b2.5 Re0.75
6.53b0.5
(b) Flow nozzle (8–72)
(8–73)
Re0.5
These relations are valid for 0.25 * b * 0.75 and 104 * Re * 107. Precise values of Cd depend on the particular design of the obstruction, and thus the manufacturer’s data should be consulted when available. For flows with high Reynolds numbers (Re ) 30,000), the value of Cd can be taken to be 0.96 for flow nozzles and 0.61 for orifices. Owing to its streamlined design, the discharge coefficients of Venturi meters are very high, ranging between 0.95 and 0.99 (the higher values are for the higher Reynolds numbers) for most flows. In the absence of specific data, we can take Cd ! 0.98 for Venturi meters. Also, the Reynolds number depends on the flow velocity, which is not known a priori. Therefore, the solution is iterative in nature when curve-fit correlations are used for Cd. The orifice meter has the simplest design and it occupies minimal space as it consists of a plate with a hole in the middle, but there are considerable variations in design (Fig. 8–57). Some orifice meters are sharp-edged, while
Force
D
21°
d
15°
(c) Venturi meter
FIGURE 8–56 Common types of obstruction meters.
Magnet Bellows
Housing
Flow
d
P1 V1 V2 > V1
FIGURE 8–57 An orifice meter and schematic showing its built-in pressure transducer and digital readout.
P2 V2
P1 > P2
Orifice
Courtesy KOBOLD Instruments, Pittsburgh, PA. www.koboldusa.com. Used by permission.
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368 FLUID MECHANICS P1 Lost pressure Pressure drop across orifice
HGL
P3
Recovered pressure
P2
Orifice meter
FIGURE 8–58 The variation of pressure along a flow section with an orifice meter as measured with piezometer tubes; the lost pressure and the pressure recovery are shown.
others are beveled or rounded. The sudden change in the flow area in orifice meters causes considerable swirl and thus significant head loss or permanent pressure loss, as shown in Fig. 8–58. In nozzle meters, the plate is replaced by a nozzle, and thus the flow in the nozzle is streamlined. As a result, the vena contracta is practically eliminated and the head loss is small. However, flow nozzle meters are more expensive than orifice meters. The Venturi meter, invented by the American engineer Clemans Herschel (1842–1930) and named by him after the Italian Giovanni Venturi (1746– 1822) for his pioneering work on conical flow sections, is the most accurate flowmeter in this group, but it is also the most expensive. Its gradual contraction and expansion prevent flow separation and swirling, and it suffers only frictional losses on the inner wall surfaces. Venturi meters cause very low head losses, as shown in Fig. 8–59, and thus they should be preferred for applications that cannot allow large pressure drops. The irreversible head loss for Venturi meters due to friction is only about 10 percent. EXAMPLE 8–10
Measuring Flow Rate with an Orifice Meter
The flow rate of methanol at 20°C (r ! 788.4 kg/m3 and m ! 5.857 . 10&4 kg/m · s) through a 4-cm-diameter pipe is to be measured with a 3-cm-diameter orifice meter equipped with a mercury manometer across the orifice place, as shown in Fig. 8–60. If the differential height of the manometer is read to be 11 cm, determine the flow rate of methanol through the pipe and the average flow velocity.
SOLUTION The flow rate of methanol is to be measured with an orifice meter. For a given pressure drop across the orifice plate, the flow rate and the average flow velocity are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The discharge coefficient of the orifice meter is Cd ! 0.61.
1.00 90
Fraction of pressure loss, %
80
FIGURE 8–59 The fraction of pressure (or head) loss for various obstruction meters. From ASME Fluid Meters. Used by permission of ASME International.
70
Orifice with flange taps
Flow nozzle
60 50 40 30 20
Short cone Venturi Long cone Venturi
10 0 0
Lo-loss tube 0.10
0.20
0.30
0.40
0.50 0.60 d/D ratio, b
0.70
0.80
0.90
1.00
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369 CHAPTER 8
Properties The density and dynamic viscosity of methanol are given to be r ! 788.4 kg/m3 and m ! 5.857 . 10&4 kg/m · s, respectively. We take the density of mercury to be 13,600 kg/m3. Analysis The diameter ratio and the throat area of the orifice are
1
2
d 3 b ! ! ! 0.75 D 4 A0 !
pd 2 p(0.03 m)2 ! ! 7.069 . 10 &4 m2 4 4
The pressure drop across the orifice plate can be expressed as
(P ! P1 & P2 ! (rHg & rmet)gh
11 cm
Then the flow rate relation for obstruction meters becomes
# V ! A 0C d
2(r Hg & r met)gh 2(r Hg/r met & 1)gh 2(P1 & P2) ! A 0C d ! A 0C d 4 4 B r(1 & b ) B r met(1 & b ) B 1 & b4
Substituting, the flow rate is determined to be
# 2(13,600/788.4 & 1)(9.81 m/s2)(0.11 m) V ! (7.069 . 10 &4 m2)(0.61) B 1 & 0.75 4 ! 3.09 . 10 &3 m3/s which is equivalent to 3.09 L/s. The average flow velocity in the pipe is determined by dividing the flow rate by the cross-sectional area of the pipe,
V!
# # V V 3.09 . 10 &3 m3/s ! ! ! 2.46 m/s 2 A c pD /4 p(0.04 m)2/4
Discussion The Reynolds number of flow through the pipe is
Re !
rVD (788.4 kg/m3)(2.46 m/s)(0.04 m) ! ! 1.32 . 105 m 5.857 . 10 &4 kg/m , s
Substituting b ! 0.75 and Re ! 1.32 . 105 into the orifice discharge coefficient relation
Cd ! 0.5959 ' 0.0312b2.1 & 0.184b8 '
91.71b2.5 Re0.75
gives Cd ! 0.601, which is very close to the assumed value of 0.61. Using this refined value of Cd, the flow rate becomes 3.04 L/s, which differs from our original result by only 1.6 percent. Therefore, it is convenient to analyze orifice meters using the recommended value of Cd ! 0.61 for the discharge coefficient, and then to verify the assumed value. If the problem is solved using an equation solver such as EES, then the problem can be formulated using the curve-fit formula for Cd (which depends on the Reynolds number), and all equations can be solved simultaneously by letting the equation solver perform the iterations as necessary.
Positive Displacement Flowmeters
When we buy gasoline for the car, we are interested in the total amount of gasoline that flows through the nozzle during the period we fill the tank rather than the flow rate of gasoline. Likewise, we care about the total
Mercury manometer
FIGURE 8–60 Schematic for the orifice meter considered in Example 8–10.
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370 FLUID MECHANICS
FIGURE 8–61 A positive displacement flowmeter with double helical three-lobe impeller design. Courtesy Flow Technology, Inc. Source: www.ftimeters.com.
amount of water or natural gas we use in our homes during a billing period. In these and many other applications, the quantity of interest is the total amount of mass or volume of a fluid that passes through a cross section of a pipe over a certain period of time rather than the instantaneous value of flow rate, and positive displacement flowmeters are well suited for such applications. There are numerous types of displacement meters, and they are based on continuous filling and discharging of the measuring chamber. They operate by trapping a certain amount of incoming fluid, displacing it to the discharge side of the meter, and counting the number of such discharge– recharge cycles to determine the total amount of fluid displaced. The clearance between the impeller and its casing must be controlled carefully to prevent leakage and thus to avoid error. Figure 8–61 shows a positive displacement flowmeter with two rotating impellers driven by the flowing liquid. Each impeller has three gear lobes, and a pulsed output signal is generated each time a lobe passes by a nonintrusive sensor. Each pulse represents a known volume of liquid that is captured in between the lobes of the impellers, and an electronic controller converts the pulses to volume units. This particular meter has a quoted accuracy of 0.1 percent, has a low pressure drop, and can be used with high- or lowviscosity liquids at temperatures up to 230°C and pressures up to 7 MPa for flow rates of up to 700 gal/min (or 50 L/s). The most widely used flowmeters to measure liquid volumes are nutating disk flowmeters, shown in Fig. 8–62. They are commonly used as water and gasoline meters. The liquid enters the nutating disk meter through the chamber (A). This causes the disk (B) to nutate or wobble and results in the rotation of a spindle (C) and the excitation of a magnet (D). This signal is transmitted through the casing of the meter to a second magnet (E). The total volume is obtained by counting the number of these signals during a discharge process. Quantities of gas flows, such as the amount of natural gas used in buildings, are commonly metered by using bellows flowmeters that displace a certain amount of gas volume during each revolution.
Turbine Flowmeters
We all know from experience that a propeller held against the wind rotates, and the rate of rotation increases as the wind velocity increases. You may
E D B
FIGURE 8–62 A nutating disk flowmeter. (a) Courtesy Badger Meter, Inc. Source: www.badgermeter.com.
A
C
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371 CHAPTER 8
(a)
(b)
also have seen that the turbine blades of wind turbines rotate rather slowly at low winds, but quite fast at high winds. These observations suggest that the flow velocity in a pipe can be measured by placing a freely rotating propeller inside a pipe section and doing the necessary calibration. Flow measurement devices that work on this principle are called turbine flowmeters or sometimes propeller flowmeters, although the latter is a misnomer since, by definition, propellers add energy to a fluid, while turbines extract energy from a fluid. A turbine flowmeter consists of a cylindrical flow section that houses a turbine (a vaned rotor) that is free to rotate, additional stationary vanes at the inlet to straighten the flow, and a sensor that generates a pulse each time a marked point on the turbine passes by to determine the rate of rotation. The rotational speed of the turbine is nearly proportional to the flow rate of the fluid. Turbine flowmeters give highly accurate results (as accurate as 0.25 percent) over a wide range of flow rates when calibrated properly for the anticipated flow conditions. Turbine flowmeters have very few blades (sometimes just two blades) when used to measure liquid flow, but several blades when used to measure gas flow to ensure adequate torque generation. The head loss caused by the turbine is very small. Turbine flowmeters have been used extensively for flow measurement since the 1940s because of their simplicity, low cost, and accuracy over a wide range of flow conditions. They are made commercially available for both liquids and gases and for pipes of practically all sizes. Turbine flowmeters are also commonly used to measure flow velocities in unconfined flows such as winds, rivers, and ocean currents. The handheld device shown in Fig. 8–63b is used to measure wind velocity.
Paddlewheel Flowmeters
Paddlewheel flowmeters are low-cost alternatives to turbine flowmeters for flows where very high accuracy is not required. In paddlewheel flowmeters, the paddlewheel (the rotor and the blades) is perpendicular to the flow, as shown in Fig. 8–64, rather than parallel as was the case with turbine
(C)
FIGURE 8–63 (a) An in-line turbine flowmeter to measure liquid flow, with flow from left to right, (b) a close-up view of the turbine blades inside the flowmeter, looking down the axis with flow into the page, and (c) a handheld turbine flowmeter to measure wind speed, measuring no flow so that the turbine blades are visible. The flowmeter in (c) also measures the air termperature for convenience. Photos (a) and (c) by John M. Cimbala. Photo (b) Courtesy Hoffer Flow Controls.
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372 FLUID MECHANICS Retainer cap Paddlewheel sensor Sensor housing
FIGURE 8–64 Paddlewheel flowmeter to measure liquid flow, with flow from left to right, and a schematic diagram of its operation.
Truseal locknut
Flow
Photo by John M. Cimbala.
flowmeters. The paddles cover only a portion of the flow cross section (typically, less than half), and thus the head loss is much smaller compared to that of turbine flowmeters, but the depth of insertion of the paddlewheel into the flow is of critical importance for accuracy. Also, no strainers are required since the paddlewheels are not susceptible to fouling. A sensor detects the passage of each of the paddlewheel blades and transmits a signal. A microprocessor then converts this rotational speed information to flow rate or integrated flow quantity.
Variable-Area Flowmeters (Rotameters)
(b)
(a)
FIGURE 8–65 Two types of variable-area flowmeters: (a) an ordinary gravity-based meter and (b) a spring-opposed meter. (a) Photo by Luke A. Cimbala and (b) Courtesy Insite, Universal Flow Monitors, Inc. Used by permission.
A simple, reliable, inexpensive, and easy-to-install flowmeter with low pressure drop and no electrical connections that gives a direct reading of flow rate for a wide range of liquids and gases is the variable-area flowmeter, also called a rotameter or floatmeter. A variable-area flowmeter consists of a vertical tapered conical transparent tube made of glass or plastic with a float inside that is free to move, as shown in Fig. 8–65. As fluid flows through the tapered tube, the float rises within the tube to a location where the float weight, drag force, and buoyancy force balance each other and the net force acting on the float is zero. The flow rate is determined by simply matching the position of the float against the graduated flow scale outside the tapered transparent tube. We know from experience that high winds knock down trees, break power lines, and blow away hats or umbrellas. This is because the drag force increases with flow velocity. The weight and the buoyancy force acting on the float are constant, but the drag force changes with flow velocity. Also, the velocity along the tapered tube decreases in the flow direction because of the increase in the cross-sectional area. There is a certain velocity that generates enough drag to balance the float weight and the buoyancy force, and the location at which this velocity occurs around the float is the location where the float settles. The degree of tapering of the tube can be made such that the vertical rise changes linearly with flow rate, and thus the tube can be calibrated linearly for flow rates. The transparent tube also allows the fluid to be seen during flow. There are numerous kinds of variable-area flowmeters. The gravity-based flowmeter discussed previously must be positioned vertically, with fluid entering from the bottom and leaving from the top. In spring-opposed
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flowmeters, the drag force is balanced by the spring force, and such flowmeters can be installed horizontally. Another type of flowmeter uses a loose-fitting piston instead of a float. The accuracy of variable-area flowmeters is typically 15 percent. Therefore, these flowmeters are not appropriate for applications that require precision measurements. However, some manufacturers quote accuracies of the order of 1 percent. Also, these meters depend on visual checking of the location of the float, and thus they cannot be used to measure the flow rate of fluids that are opaque or dirty, or fluids that coat the float since such fluids block visual access. Finally, glass tubes are prone to breakage and thus they pose a safety hazard if toxic fluids are handled. In such applications, variable-area flowmeters should be installed at locations with minimum traffic.
Ultrasonic Flowmeters
It is a common observation that when a stone is dropped into calm water, the waves that are generated spread out as concentric circles uniformly in all directions. But when a stone is thrown into flowing water such as a river, the waves move much faster in the flow direction (the wave and flow velocities are added since they are in the same direction) compared to the waves moving in the upstream direction (the wave and flow velocities are subtracted since they are in opposite directions). As a result, the waves appear spread out downstream while they appear tightly packed upstream. The difference between the number of waves in the upstream and downstream parts of the flow per unit length is proportional to the flow velocity, and this suggests that flow velocity can be measured by comparing the propagation of waves in the forward and backward directions to flow. Ultrasonic flowmeters operate on this principle, using sound waves in the ultrasonic range (typically at a frequency of 1 MHz). Ultrasonic (or acoustic) flowmeters operate by generating sound waves with a transducer and measuring the propagation of those waves through a flowing fluid. There are two basic kinds of ultrasonic flowmeters: transit time and Doppler-effect (or frequency shift) flowmeters. The transit time flowmeter transmits sound waves in the upstream and downstream directions and measures the difference in travel time. A typical transit time ultrasonic meter is shown schematically in Fig. 8–66. It involves two transducers that alternately transmit and receive ultrasonic waves, one in the direction of flow and the other in the opposite direction. The travel time for each direction can be measured accurately, and the difference in the travel time can be calculated. The average flow velocity V in the pipe is proportional to this travel time difference (t, and can be determined from V ! KL (t
(8–74)
where L is the distance between the transducers and K is a constant.
Doppler-Effect Ultrasonic Flowmeters
You have probably noticed that when a fast-moving car approaches with its horn blowing, the tone of the high-pitched sound of the horn drops to a lower pitch as the car passes by. This is due to the sonic waves being compressed in front of the car and being spread out behind it. This shift in frequency is called the Doppler effect, and it forms the basis for the operation of most ultrasonic flowmeters.
Top view
Flow A
Reflect-mode configuration
B
FIGURE 8–66 The operation of a transit time ultrasonic flowmeter equipped with two transducers, www.flocat.com.
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374 FLUID MECHANICS Transmitting element
Reflectors
Receiving element
Flow direction
FIGURE 8–67 The operation of a Doppler-effect ultrasonic flowmeter equipped with a transducer pressed on the outer surface of a pipe.
FIGURE 8–68 Ultrasonic flowmeters enable one to measure flow velocity without even contacting the fluid by simply pressing a transducer on the outer surface of the pipe. Photo by J. Matthew Deepe.
Doppler-effect ultrasonic flowmeters measure the average flow velocity along the sonic path. This is done by clamping a piezoelectric transducer on the outside surface of a pipe (or pressing the transducer against the pipe for handheld units). The transducer transmits a sound wave at a fixed frequency through the pipe wall and into the flowing liquid. The waves reflected by impurities, such as suspended solid particles or entrained gas bubbles, are relayed to a receiving transducer. The change in the frequency of the reflected waves is proportional to the flow velocity, and a microprocessor determines the flow velocity by comparing the frequency shift between the transmitted and reflected signals (Figs. 8–67 and 8–68). The flow rate and the total amount of flow can also be determined using the measured velocity by properly configuring the flowmeter for the given pipe and flow conditions. The operation of ultrasonic flowmeters depends on the ultrasound waves being reflected off discontinuities in density. Ordinary ultrasonic flowmeters require the liquid to contain impurities in concentrations greater than 25 parts per million (ppm) in sizes greater than at least 30 4m. But advanced ultrasonic units can also measure the velocity of clean liquids by sensing the waves reflected off turbulent swirls and eddies in the flow stream, provided that they are installed at locations where such disturbances are nonsymmetrical and at a high level, such as a flow section just downstream of a 90° elbow. Ultrasonic flowmeters have the following advantages: • They are easy and quick to install by clamping them on the outside of pipes of 0.6 cm to over 3 m in diameter, and even on open channels. • They are nonintrusive. Since the meters clamp on, there is no need to stop operation and drill holes into piping, and no production downtime. • There is no pressure drop since the meters do not interfere with the flow.
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• Since there is no direct contact with the fluid, there is no danger of corrosion or clogging. • They are suitable for a wide range of fluids from toxic chemicals to slurries to clean liquids, for permanent or temporary flow measurement. • There are no moving parts, and thus the meters provide reliable and maintenance-free operation. • They can also measure flow quantities in reverse flow. • The quoted accuracies are 1 to 2 percent. Ultrasonic flowmeters are noninvasive devices, and the ultrasonic transducers can effectively transmit signals through polyvinyl chloride (PVC), steel, iron, and glass pipe walls. However, coated pipes and concrete pipes are not suitable for this measurement technique since they absorb ultrasonic waves.
Electromagnetic Flowmeters
It has been known since Faraday’s experiments in the 1830s that when a conductor is moved in a magnetic field, an electromotive force develops across that conductor as a result of magnetic induction. Faraday’s law states that the voltage induced across any conductor as it moves at right angles through a magnetic field is proportional to the velocity of that conductor. This suggests that we may be able to determine flow velocity by replacing the solid conductor by a conducting fluid, and electromagnetic flowmeters do just that. Electromagnetic flowmeters have been in use since the mid1950s, and they come in various designs such as full-flow and insertion types. A full-flow electromagnetic flowmeter is a nonintrusive device that consists of a magnetic coil that encircles the pipe, and two electrodes drilled into the pipe along a diameter flush with the inner surface of the pipe so that the electrodes are in contact with the fluid but do not interfere with the flow and thus do not cause any head loss (Fig. 8–69a). The electrodes are connected to a voltmeter. The coils generate a magnetic field when subjected to electric current, and the voltmeter measures the electric potential
E Flow
E
Flow
Flow Electrodes
(a) Full-flow electromagnetic flowmeter
(b) Insertion electromagnetic flowmeter
FIGURE 8–69 (a) Full-flow and (b) insertion electromagnetic flowmeters, www.flocat.com.
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difference between the electrodes. This potential difference is proportional to the flow velocity of the conducting fluid, and thus the flow velocity can be calculated by relating it to the voltage generated. Insertion electromagnetic flowmeters operate similarly, but the magnetic field is confined within a flow channel at the tip of a rod inserted into the flow, as shown in Fig. 8–69b. Electromagnetic flowmeters are well-suited for measuring flow velocities of liquid metals such as mercury, sodium, and potassium that are used in some nuclear reactors. They can also be used for liquids that are poor conductors, such as water, provided that they contain an adequate amount of charged particles. Blood and seawater, for example, contain sufficient amounts of ions, and thus electromagnetic flowmeters can be used to measure their flow rates. Electromagnetic flowmeters can also be used to measure the flow rates of chemicals, pharmaceuticals, cosmetics, corrosive liquids, beverages, fertilizers, and numerous slurries and sludges, provided that the substances have high enough electrical conductivities. Electromagnetic flowmeters are not suitable for use with distilled or deionized water. Electromagnetic flowmeters measure flow velocity indirectly, and thus careful calibration is important during installation. Their use is limited by their relatively high cost, power consumption, and the restrictions on the types of suitable fluids with which they can be used.
Vortex Flowmeters
Bluff body (strut)
Receiving transducer
Flow
Vortex swirl
Transmitting transducer
FIGURE 8–70 The operation of a vortex flowmeter, www.flocat.com.
You have probably noticed that when a flow stream such as a river encounters an obstruction such as a rock, the fluid separates and moves around the rock. But the presence of the rock is felt for some distance downstream via the swirls generated by it. Most flows encountered in practice are turbulent, and a disk or a short cylinder placed in the flow coaxially sheds vortices (see also Chap. 4). It is observed that these vortices are shed periodically, and the shedding frequency is proportional to the average flow velocity. This suggests that the flow rate can be determined by generating vortices in the flow by placing an obstruction along the flow and measuring the shedding frequency. The flow measurement devices that work on this principle are called vortex flowmeters. The Strouhal number, defined as St ! fd/V, where f is the vortex shedding frequency, d is the characteristic diameter or width of the obstruction, and V is the velocity of the flow impinging on the obstruction, also remains constant in this case, provided that the flow velocity is high enough. A vortex flowmeter consists of a sharp-edged bluff body (strut) placed in the flow that serves as the vortex generator, and a detector (such as a pressure transducer that records the oscillation in pressure) placed a short distance downstream on the inner surface of the casing to measure the shedding frequency. The detector can be an ultrasonic, electronic, or fiber-optic sensor that monitors the changes in the vortex pattern and transmits a pulsating output signal (Fig. 8–70). A microprocessor then uses the frequency information to calculate and display the flow velocity or flow rate. The frequency of vortex shedding is proportional to the average velocity over a wide range of Reynolds numbers, and vortex flowmeters operate reliably and accurately at Reynolds numbers from 104 to 107.
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The vortex flowmeter has the advantage that it has no moving parts and thus is inherently reliable, versatile, and very accurate (usually 11 percent over a wide range of flow rates), but it obstructs flow and thus causes considerable head loss.
Thermal (Hot-Wire and Hot-Film) Anemometers
Thermal anemometers were introduced in the late 1950s and have been in common use since then in fluid research facilities and labs. As the name implies, thermal anemometers involve an electrically heated sensor, as shown in Fig. 8–71, and utilize a thermal effect to measure flow velocity. Thermal anemometers have extremely small sensors, and thus they can be used to measure the instantaneous velocity at any point in the flow without appreciably disturbing the flow. They can take thousands of velocity measurements per second with excellent spatial and temporal resolution, and thus they can be used to study the details of fluctuations in turbulent flow. They can measure velocities in liquids and gases accurately over a wide range—from a few centimeters to over a hundred meters per second. A thermal anemometer is called a hot-wire anemometer if the sensing element is a wire, and a hot-film anemometer if the sensor is a thin metallic film (less than 0.1 4m thick) mounted usually on a relatively thick ceramic support having a diameter of about 50 4m. The hot-wire anemometer is characterized by its very small sensor wire—usually a few microns in diameter and a couple of millimeters in length. The sensor is usually made of platinum, tungsten, or platinum–iridium alloys, and it is attached to the probe through holders. The fine wire sensor of a hot-wire anemometer is very fragile because of its small size and can easily break if the liquid or gas contains excessive amounts of contaminants or particulate matter. This is especially of consequence at high velocities. In such cases, the more rugged hot-film probes should be used. But the sensor of the hot-film probe is larger, has significantly lower frequency response, and interferes more with the flow; thus it is not always suitable for studying the fine details of turbulent flow. The operating principle of a constant-temperature anemometer (CTA), which is the most common type and is shown schematically in Fig. 8–72, is as follows: the sensor is electrically heated to a specified temperature (typically about 200°C). The sensor tends to cool as it loses heat to the surrounding flowing fluid, but electronic controls maintain the sensor at a constant
Electric current I
Flow velocity V
Sensor (a thin wire approximately1 mm long with a diameter of 5 mm)
Wire support
FIGURE 8–71 The electrically heated sensor and its support of a hot-wire probe.
Signal conditioner
CTA
Sensor Probe
Flow
Bridge
Servo loop
Filter
Gain
Connector box and computer
FIGURE 8–72 Schematic of a thermal anemometer system.
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FIGURE 8–73 Thermal anemometer probes with single, double, and triple sensors to measure (a) one-, (b) two-, and (c) three-dimensional velocity components simultaneously.
(a)
(b)
(c)
temperature by varying the electric current (which is done by varying the voltage) as needed. The higher the flow velocity, the higher the rate of heat transfer from the sensor, and thus the larger the voltage that needs to be applied across the sensor to maintain it at constant temperature. There is a close correlation between the flow velocity and voltage, and the flow velocity can be determined by measuring the voltage applied by an amplifier or the electric current passing through the sensor. The sensor is maintained at a constant temperature during operation, and thus its thermal energy content remains constant. The conservation of . energy principle requires that the electrical Joule heating Welect ! I 2Rw ! E 2/Rw. of the sensor must be equal to the total rate of heat loss from the sensor Qtotal, which consists of convection heat transfer since conduction to the wire supports and radiation to the surrounding surfaces are small and can be disregarded. Using proper relations for forced convection, the energy balance can be expressed by King’s law as E 2 ! a ' bV n
(8–75)
where E is the voltage, and the values of the constants a, b, and n are calibrated for a given probe. Once the voltage is measured, this relation gives the flow velocity V directly. Most hot-wire sensors have a diameter of 5 mm and a length of approximately 1 mm and are made of tungsten. The wire is spot-welded to needleshaped prongs embedded in a probe body, which is connected to the anemometer electronics. Thermal anemometers can be used to measure twoor three-dimensional velocity components simultaneously by using probes with two or three sensors, respectively (Fig. 8–73). When selecting probes, consideration should be given to the type and the contamination level of the fluid, the number of velocity components to be measured, the required spatial and temporal resolution, and the location of measurement.
Laser Doppler Velocimetry
Laser Doppler velocimetry (LDV), also called laser velocimetry (LV) or laser Doppler anemometry (LDA), is an optical technique to measure flow velocity at any desired point without disturbing the flow. Unlike thermal anemometry, LDV involves no probes or wires inserted into the flow, and thus it is a nonintrusive method. Like thermal anemometry, it can accurately measure velocity at a very small volume, and thus it can also be used to study the details of flow at a locality, including turbulent fluctuations, and it can be traversed through the entire flow field without intrusion. The LDV technique was developed in the mid-1960s and has found widespread acceptance because of the high accuracy it provides for both gas and
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liquid flows; the high spatial resolution it offers; and, in recent years, its ability to measure all three velocity components. Its drawbacks are the relatively high cost; the requirement for sufficient transparency between the laser source, the target location in the flow, and the photodetector; and the requirement for careful alignment of emitted and reflected beams for accuracy. The latter drawback is eliminated for the case of a fiber-optic LDV system, since it is aligned at the factory. The operating principle of LDV is based on sending a highly coherent monochromatic (all waves are in phase and at the same wavelength) light beam toward the target, collecting the light reflected by small particles in the target area, determining the change in frequency of the reflected radiation due to the Doppler effect, and relating this frequency shift to the flow velocity of the fluid at the target area. LDV systems are available in many different configurations. A basic dualbeam LDV system to measure a single velocity component is shown in Fig. 8–74. In the heart of all LDV systems is a laser power source, which is usually a helium–neon or argon-ion laser with a power output of 10 mW to 20 W. Lasers are preferred over other light sources since laser beams are highly coherent and highly focused. The helium–neon laser, for example, emits radiation at a wavelength of 0.6328 4m, which is in the reddish-orange color range. The laser beam is first split into two parallel beams of equal intensity by a half-silvered mirror called a beam splitter. Both beams then pass through a converging lens that focuses the beams at a point in the flow (the target). The small fluid volume where the two beams intersect is the region where the velocity is measured and is called the measurement volume or the focal volume. The measurement volume resembles an ellipsoid, typically of 0.1 mm diameter and 0.5 mm in length. The laser light is scattered by particles passing through this measurement volume, and the light scattered in a certain direction is collected by a receiving lens and is passed through a photodetector that converts the fluctuations in light intensity into fluctuations in a voltage signal. Finally, a signal processor determines the frequency of the voltage signal and thus the velocity of the flow. The waves of the two laser beams that cross in the measurement volume are shown schematically in Fig. 8–75. The waves of the two beams interfere in the measurement volume, creating a bright fringe where they are in phase and thus support each other, and creating a dark fringe where they are out of phase and thus cancel each other. The bright and dark fringes form lines
Photodetector Receiving lens Beam splitter
Sending lens
Laser
V a Measurement volume
Mirror Bragg cell
FIGURE 8–74 A dual-beam LDV system in forward scatter mode.
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380 FLUID MECHANICS
Fringe lines
a
parallel to the midplane between the two incident laser beams. Using trigonometry, the spacing s between the fringe lines, which can be viewed as the wavelength of fringes, can be shown to be s ! l/[2 sin(a/2)], where l is the wavelength of the laser beam and a is the angle between the two laser beams. When a particle traverses these fringe lines at velocity V, the frequency of the scattered fringe lines is
s
l
Laser beams
Measurement volume
V
f!
Fringe lines
FIGURE 8–75 Fringes that form as a result of the interference at the intersection of two laser beams of an LDV system (lines represent peaks of waves). The top diagram is a close-up view of two fringes.
5
V 2V sin(a/2) ! s l
(8–76)
This fundamental relation shows the flow velocity to be proportional to the frequency and is known as the LDV equation. As a particle passes through the measurement volume, the reflected light is bright, then dark, then bright, etc., because of the fringe pattern, and the flow velocity is determined by measuring the frequency of the reflected light. The velocity profile at a cross section of a pipe can be obtained by mapping the flow across the pipe (Fig. 8–76). The LDV method obviously depends on the presence of scattered fringe lines, and thus the flow must contain a sufficient amount of small particles called seeds or seeding particles. These particles must be small enough to follow the flow closely so that the particle velocity is equal to the flow velocity, but large enough (relative to the wavelength of the laser light) to scatter an adequate amount of light. Particles with a diameter of 1 4m usually serve the purpose well. Some fluids such as tap water naturally contain an adequate amount of such particles, and no seeding is necessary. Gases such as air are commonly seeded with smoke or with particles made of latex, oil, or other materials. By using three laser beam pairs at different wavelengths, the LDV system is also used to obtain all three velocity components at any point in the flow.
(m/s)
Particle Image Velocimetry
4
3
2
1
–80
–60
–40 x (mm)
–20
FIGURE 8–76 A time-averaged velocity profile in turbulent pipe flow obtained by an LDV system. Courtesy Dantec Dynamics, Inc. www.dantecmt.com. Used by permission.
0
Particle image velocimetry (PIV) is a double-pulsed laser technique used to measure the instantaneous velocity distribution in a plane of flow by photographically determining the displacement of particles in the plane during a very short time interval. Unlike methods like hot-wire anemometry and LDV that measure velocity at a point, PIV provides velocity values simultaneously throughout an entire cross section, and thus it is a whole-field technique. PIV combines the accuracy of LDV with the capability of flow visualization and provides instantaneous flow field mapping. The entire instantaneous velocity profile at a cross section of pipe, for example, can be obtained with a single PIV measurement. A PIV system can be viewed as a camera that can take a snapshot of velocity distribution at any desired plane in a flow. Ordinary flow visualization gives a qualitative picture of the details of flow. PIV also provides an accurate quantitative description of various flow quantities such as the velocity field, and thus the capability to analyze the flow numerically using the velocity data provided. Because of its whole-field capability, PIV is also used to validate computational fluid dynamics (CFD) codes (Chap. 15). The PIV technique has been used since the mid-1980s, and its use and capabilities have grown in recent years with improvements in frame grabber
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and charge-coupled device (CCD) camera technologies. The accuracy, flexibility, and versatility of PIV systems with their ability to capture whole-field images with submicrosecond exposure time have made them extremely valuable tools in the study of supersonic flows, explosions, flame propagation, bubble growth and collapse, turbulence, and unsteady flow. The PIV technique for velocity measurement consists of two main steps: visualization and image processing. The first step is to seed the flow with suitable particles in order to trace the fluid motion. Then a pulse of laser light sheet illuminates a thin slice of the flow field at the desired plane, and the positions of particles in that plane are determined by detecting the light scattered by particles on a digital video or photographic camera positioned at right angles to the light sheet (Fig. 8–77). After a very short time period (t (typically in 4s), the particles are illuminated again by a second pulse of laser light sheet, and their new positions are recorded. Using the information on these two superimposed camera images, the particle displacements (s are determined for all particles, and the magnitude of velocity of the particles in the plane of the laser light sheet is determined from (s/(t. The direction of motion of the particles is also determined from the two positions, so that two components of velocity in the plane are calculated. The built-in algorithms of PIV systems determine the velocities at thousands of area elements called interrogation regions throughout the entire plane and display the velocity field on the computer monitor in any desired form (Fig. 8–78). The PIV technique relies on the laser light scattered by particles, and thus the flow must be seeded if necessary with particles, also called markers, in order to obtain an adequate reflected signal. Seed particles must be able to follow the pathlines in the flow for their motion to be representative of the
Beam dump Video camera
Computer Seeded flow
Synchronizer Pulser
Sheet-forming optics
Pulsed Nd:YAG laser
FIGURE 8–77 A PIV system to study flame stabilization. Courtesy of TSI Incorporated (www.tsi.com). Used by permission.
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382 FLUID MECHANICS
FIGURE 8–78 Instantaneous velocity field in the wake region of a car as measured by a PIV system in a wind tunnel. The velocity vectors are superimposed on a contour plot of pressure. The interface between two adjacent grayscale levels is an isobar. Courtesy Dantec Dynamics, Inc. www.dantecmt.com.
Light-guide delivery of laser sheet Stereoscopic camera setup
Main flow y x Jet flow Jet trajectory
Field of view
FIGURE 8–79 A three-dimensional PIV system set up to study the mixing of an air jet with cross duct flow. Courtesy TSI Incorporated (www.tsi.com). Used by permission.
flow, and this requires the particle density to be equal to the fluid density (so that they are neutrally buoyant) or the particles to be so small (typically 4m-sized) that their movement relative to the fluid is insignificant. A variety of such particles is available to seed gas or liquid flow. Very small particles must be used in high-speed flows. Silicon carbide particles (mean diameter of 1.5 4m) are suitable for both liquid and gas flow, titanium dioxide particles (mean diameter of 0.2 4m) are usually used for gas flow and are suitable for high-temperature applications, and polystyrene latex particles (nominal diameter of 1.0 4m) are suitable for low-temperature applications. Metallic-coated particles (mean diameter of 9.0 4m) are also used to seed water flows for LDV measurements because of their high reflectivity. Gas bubbles as well as droplets of some liquids such as olive oil or silicon oil are also used as seeding particles after they are atomized to 4m-sized spheres. A variety of laser light sources such as argon, copper vapor, and Nd:YAG can be used with PIV systems, depending on the requirements for pulse duration, power, and time between pulses. Nd:YAG lasers are commonly used in PIV systems over a wide range of applications. A beam delivery system such as a light arm or a fiber-optic system is used to generate and deliver a high-energy pulsed laser sheet at a specified thickness. With PIV, other flow properties such as vorticity and strain rates can also be obtained, and the details of turbulence can be studied. Recent advances in PIV technology have made it possible to obtain three-dimensional velocity profiles at a cross section of a flow using two cameras (Fig. 8–79). This is done by recording the images of the target plane simultaneously by both cameras at different angles, processing the information to produce two separate two-dimensional velocity maps, and combining these two maps to generate the instantaneous three-dimensional velocity field.
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APPLICATION SPOTLIGHT
■
How Orifice Plate Flowmeters Work, or Do Not Work
Guest Author: Lorenz Sigurdson, Vortex Fluid Dynamics Lab, University of Alberta The Bernoulli equation is the most beloved of all fluid mechanical equations because it is a scalar equation and has a vast range of applications. One very valuable use is in the development of Bernoulli obstruction theory. This theory allows an estimate of the flow velocity from the measured pressure drop between locations upstream and downstream of an obstruction in a pipe flow. The volume flow rate can be calculated by using the Bernoulli equation, conservation of mass, and the obstruction geometry. The cheapest obstruction to produce is a plate with a circular orifice in it. There are hundreds of thousands of orifice plate flowmeters in use in North America. It is the accepted international standard of measurement of volume flow rates. The accuracy can become very important in industries such as natural gas pipelining where the commodity is bought and sold based on measurements from these meters. Some pipes carry more than a million dollars per hour of natural gas. For practical purposes, meter calibration is required because, although the pipe and orifice diameter may be known, the flow separates from the lip of the orifice and creates a flow tube narrower than the orifice diameter. The flow is accelerated through this vena contracta. Figure 8–80 shows the flow downstream of the orifice visualized by using a smoke-wire to introduce streaklines in a transparent flowmeter. The calibration assumes that there is no pulsation in the pipe flow. However, this is not the case in practice if there is a reciprocating compressor in the pipeline, or a loose flapping valve. Figure 8–81 shows what can happen to the vena contracta in this circumstance, if the frequency of the pulsation is near a resonance frequency of the turbulent flow structures. The vena contracta diameter is reduced. Stop reading and ask yourself, “Will this cause a flow rate underprediction or overprediction?” Conservation of mass and the narrower vena contracta mean a higher average velocity there than without pulsation. The Bernoulli equation says that the pressure will be lower there as a result, meaning a larger pressure drop and an overprediction. Errors as high as 40 percent have been found at high pulsation levels. For the natural gas pipeline mentioned, that could mean paying (or earning) $400,000 too much per hour! Characteristic instabilities that have previously been found in shear flows, jet flows, and reattaching flows (Sigurdson, 1995; Sigurdson and Chapple, 1997) also exist downstream of the orifice plate. Thankfully, meter installation designers can now avoid the dangerously resonant pulsation frequencies associated with these instabilities, thereby minimizing flowmeter error. References Sigurdson, L. W., “The Structure and Control of a Turbulent Reattaching Flow,” J. Fluid Mechanics, 298, pp. 139–165, 1995. Sigurdson, L. W., and Chapple, D., “Visualization of Acoustically Pulsated Flow through an Orifice Plate Flow Meter,” Proc. 1st Pacific Image Processing and Flow Visualization Conf., Honolulu, HI, February 23–26, 1997. Sigurdson, L. W., and Chapple, D., “A Turbulent Mechanism for Pulsation— Induced Orifice Plate Flow Meter Error,” Proc. 13th Australasian Fluid Mechanics Conf., December 13–18, 1998, Monash U., Melbourne, Australia, Thompson, M.C., and Hourigan, K., eds., 1, pp. 67–70, 1998.
Re!9000 No Pulsation Dv ! 59mm
FIGURE 8–80 Smoke-wire streakline photograph of orifice plate flowmeter with no pulsation present, Reynolds number ! 9000. Dv indicates the estimated vena contracta diameter. A hot-wire probe can be seen along the pipe centerline. From Sigurdson and Chapple (1998).
Dv!57mm
Re!9000 Stj!0.42 u7/Uv!13% (CD!13%
68mm
FIGURE 8–81 Smoke-wire streakline photograph of orifice plate flowmeter with pulsation present, showing a large effect. The meter is in error by 13%. Reynolds number ! 9000. The vena contracta diameter Dv is reduced from the no pulsation case of Fig. 8–80. From Sigurdson and Chapple (1998).
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SUMMARY In internal flow, a pipe is completely filled with a fluid. Laminar flow is characterized by smooth streamlines and highly ordered motion, and turbulent flow is characterized by velocity fluctuations and highly disordered motion. The Reynolds number is defined as VavgD rVavgD Inertial forces Re ! ! ! m n Viscous forces Under most practical conditions, the flow in a pipe is laminar at Re * 2300, turbulent at Re ) 4000, and transitional in between. The region of the flow in which the effects of the viscous shearing forces are felt is called the velocity boundary layer. The region from the pipe inlet to the point at which the boundary layer merges at the centerline is called the hydrodynamic entrance region, and the length of this region is called the hydrodynamic entry length Lh. It is given by L h, laminar " 0.05 Re D and L h, turbulent " 10D The friction coefficient in the fully developed flow region remains constant. The maximum and average velocities in fully developed laminar flow in a circular pipe are u max ! 2Vavg
and
(PD2 Vavg ! 32mL
The volume flow rate and the pressure drop for laminar flow in a horizontal pipe are # (PpD4 V ! Vavg A c ! 128mL
and (P !
D2
((P & rgL sin u)D2 and 32mL
# ((P & rgL sin u)pD V! 128mL
4
The pressure loss and head loss for all types of internal flows (laminar or turbulent, in circular or noncircular pipes, smooth or rough surfaces) are expressed as (PL ! f
L rV 2 D 2
and
hL !
1 2f
e/D 2.51 b ' 3.7 Re2f
! &2.0 loga
The plot of this formula is known as the Moody chart. The design and analysis of piping systems involve the determination of the head loss, flow rate, or the pipe diameter. Tedious iterations in these calculations can be avoided by the approximate Swamee–Jain formulas expressed as # V 2L e nD 0.9 &2 hL ! 1.07 5 eln c ' 4.62 a # b df 3.7D gD V 10 &6 * e/D * 10 &2 3000 * Re * 3 . 108
# gD5hL 0.5 3.17n 2L 0.5 e V ! &0.965a b lnc 'a b d L 3.7D gD3hL
(PL L V2 !f rg D 2g
where rV 2/2 is the dynamic pressure and the dimensionless quantity f is the friction factor. For fully developed laminar flow in a circular pipe, the friction factor is f ! 64/Re. For noncircular pipes, the diameter in the previous relations is replaced by the hydraulic diameter defined as Dh ! 4Ac /p, where Ac is the cross-sectional area of the pipe and p is its wetted perimeter.
Re ) 2000
# # LV 2 4.75 L 5.2 0.04 b ' nV 9.4 a b d ghL ghL
D ! 0.66ce 1.25 a
10 &6 * e/D * 10 &2 5000 * Re * 3 . 108
32mLVavg
These results for horizontal pipes can also be used for inclined pipes provided that (P is replaced by (P & rgL sin u, Vavg !
In fully developed turbulent flow, the friction factor depends on the Reynolds number and the relative roughness e/D. The friction factor in turbulent flow is given by the Colebrook equation, expressed as
The losses that occur in piping components such as fittings, valves, bends, elbows, tees, inlets, exits, enlargements, and contractions are called minor losses. The minor losses are usually expressed in terms of the loss coefficient KL. The head loss for a component is determined from hL ! K L
V2 2g
When all the loss coefficients are available, the total head loss in a piping system is determined from V 2j L i V 2i h L, total ! hL, major ' hL, minor ! a f i ' a K L, j D i 2g 2g i j If the entire piping system has a constant diameter, the total head loss reduces to h L, total ! af
L V2 ' a K Lb D 2g
The analysis of a piping system is based on two simple principles: (1) The conservation of mass throughout the system must be satisfied and (2) the pressure drop between two points must be the same for all paths between the two points.
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385 CHAPTER 8
When the pipes are connected in series, the flow rate through the entire system remains constant regardless of the diameters of the individual pipes. For a pipe that branches out into two (or more) parallel pipes and then rejoins at a junction downstream, the total flow rate is the sum of the flow rates in the individual pipes but the head loss in each branch is the same. When a piping system involves a pump and/or turbine, the steady-flow energy equation is expressed as V 21 P1 ' a1 ' z 1 ' hpump, u rg 2g !
P2 V 22 ' a 2 ' z 2 ' hturbine, e ' hL rg 2g
When the useful pump head hpump, u is known, the mechanical power that needs to be supplied by the pump to the fluid and the electric power consumed by the motor of the pump for a specified flow rate are determined from # # rV ghpump, u rV ghpump, u # # and W elect ! W pump, shaft ! h pump h pump–motor
where hpump–motor is the efficiency of the pump–motor combination, which is the product of the pump and the motor efficiencies. . The plot of the head loss versus the flow rate V is called the system curve. The head produced by a pump is. not a constant, and the curves of hpump, u and hpump versus V are called the characteristic curves. A pump installed in a piping system operates at the operating point, which is the point of intersection of the system curve and the characteristic curve. Flow measurement techniques and devices can be considered in three major categories: (1) volume (or mass) flow rate measurement techniques and devices such as obstruction flowmeters, turbine meters, positive displacement flowmeters, rotameters, and ultrasonic meters; (2) point velocity measurement techniques such as the Pitot-static probes, hotwires, and LDV; and (3) whole-field velocity measurement techniques such as PIV. The emphasis in this chapter has been on flow through pipes. A detailed treatment of numerous types of pumps and turbines, including their operation principles and performance parameters, is given in Chap. 14.
REFERENCES AND SUGGESTED READING 1. H. S. Bean (ed.). Fluid Meters: Their Theory and Applications, 6th ed. New York: American Society of Mechanical Engineers, 1971. 2. M. S. Bhatti and R. K. Shah. “Turbulent and Transition Flow Convective Heat Transfer in Ducts.” In Handbook of Single-Phase Convective Heat Transfer, ed. S. Kakaç, R. K. Shah, and W. Aung. New York: Wiley Interscience, 1987. 3. C. F. Colebrook. “Turbulent Flow in Pipes, with Particular Reference to the Transition between the Smooth and Rough Pipe Laws,” Journal of the Institute of Civil Engineers London. 11 (1939), pp. 133–156. 4. C. T. Crowe, J. A. Roberson, and D. F. Elger. Engineering Fluid Mechanics, 7th ed. New York: Wiley, 2001. 5. F. Durst, A. Melling, and J. H. Whitelaw. Principles and Practice of Laser-Doppler Anemometry, 2nd ed. New York: Academic, 1981. 6. R. W. Fox and A. T. McDonald. Introduction to Fluid Mechanics, 5th ed. New York: Wiley, 1999. 7. Fundamentals of Orifice Meter Measurement. Houston, TX: Daniel Measurement and Control, 1997. 8. S. E. Haaland. “Simple and Explicit Formulas for the Friction Factor in Turbulent Pipe Flow,” Journal of Fluids Engineering, March 1983, pp. 89–90. 9. I. E. Idelchik. Handbook of Hydraulic Resistance, 3rd ed. Boca Raton, FL: CRC Press, 1993.
10. W. M. Kays and M. E. Crawford. Convective Heat and Mass Transfer, 3rd ed. New York: McGraw-Hill, 1993. 11. R. W. Miller. Flow Measurement Engineering Handbook, 3rd ed. New York: McGraw-Hill, 1997. 12. L. F. Moody. “Friction Factors for Pipe Flows,” Transactions of the ASME 66 (1944), pp. 671–684. 13. B. R. Munson, D. F. Young, and T. Okiishi. Fundamentals of Fluid Mechanics, 4th ed. New York: Wiley, 2002. 14. O. Reynolds. “On the Experimental Investigation of the Circumstances Which Determine Whether the Motion of Water Shall Be Direct or Sinuous, and the Law of Resistance in Parallel Channels.” Philosophical Transactions of the Royal Society of London, 174 (1883), pp. 935–982. 15. H. Schlichting. Boundary Layer Theory, 7th ed. New York: McGraw-Hill, 1979. 16. R. K. Shah and M. S. Bhatti. “Laminar Convective Heat Transfer in Ducts.” In Handbook of Single-Phase Convective Heat Transfer, ed. S. Kakaç, R. K. Shah, and W. Aung. New York: Wiley Interscience, 1987. 17. P. L. Skousen. Valve Handbook. New York: McGraw-Hill, 1998. 18. P. K. Swamee and A. K. Jain. “Explicit Equations for Pipe-Flow Problems,” Journal of the Hydraulics Division. ASCE 102, no. HY5 (May 1976), pp. 657–664.
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19. G. Vass. “Ultrasonic Flowmeter Basics,” Sensors, 14, no. 10 (1997).
21. F. M. White. Fluid Mechanics, 5th ed. New York: McGraw-Hill, 2003.
20. A. J. Wheeler and A. R. Ganji. Introduction to Engineering Experimentation. Englewood Cliffs, NJ: Prentice-Hall, 1996.
22. W. Zhi-qing. “Study on Correction Coefficients of Laminar and Turbulent Entrance Region Effects in Round Pipes,” Applied Mathematical Mechanics, 3 (1982), p. 433.
PROBLEMS* Laminar and Turbulent Flow 8–1C
Why are liquids usually transported in circular pipes?
8–2C What is the physical significance of the Reynolds number? How is it defined for (a) flow in a circular pipe of inner diameter D and (b) flow in a rectangular duct of cross section a . b?
D
a
8–10C Consider laminar flow in a circular pipe. Will the wall shear stress tw be higher near the inlet of the pipe or near the exit? Why? What would your response be if the flow were turbulent? 8–11C How does surface roughness affect the pressure drop in a pipe if the flow is turbulent? What would your response be if the flow were laminar?
Fully Developed Flow in Pipes b
FIGURE P8–2C 8–3C Consider a person walking first in air and then in water at the same speed. For which motion will the Reynolds number be higher? 8–4C Show that the Reynolds number for flow in a circular . pipe of diameter D can be expressed as Re ! 4m /(pDm). 8–5C Which fluid at room temperature requires a larger pump to flow at a specified velocity in a given pipe: water or engine oil? Why? 8–6C What is the generally accepted value of the Reynolds number above which the flow in smooth pipes is turbulent? 8–7C Consider the flow of air and water in pipes of the same diameter, at the same temperature, and at the same mean velocity. Which flow is more likely to be turbulent? Why? 8–8C What is hydraulic diameter? How is it defined? What is it equal to for a circular pipe of diameter D? 8–9C How is the hydrodynamic entry length defined for flow in a pipe? Is the entry length longer in laminar or turbulent flow? * Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
8–12C How does the wall shear stress tw vary along the flow direction in the fully developed region in (a) laminar flow and (b) turbulent flow? 8–13C What fluid property is responsible for the development of the velocity boundary layer? For what kinds of fluids will there be no velocity boundary layer in a pipe? 8–14C In the fully developed region of flow in a circular pipe, will the velocity profile change in the flow direction? 8–15C How is the friction factor for flow in a pipe related to the pressure loss? How is the pressure loss related to the pumping power requirement for a given mass flow rate? 8–16C Someone claims that the shear stress at the center of a circular pipe during fully developed laminar flow is zero. Do you agree with this claim? Explain. 8–17C Someone claims that in fully developed turbulent flow in a pipe, the shear stress is a maximum at the pipe surface. Do you agree with this claim? Explain. 8–18C Consider fully developed flow in a circular pipe with negligible entrance effects. If the length of the pipe is doubled, the head loss will (a) double, (b) more than double, (c) less than double, (d) reduce by half, or (e) remain constant. 8–19C Someone claims that the volume flow rate in a circular pipe with laminar flow can be determined by measuring the velocity at the centerline in the fully developed region, multiplying it by the cross-sectional area, and dividing the result by 2. Do you agree? Explain. 8–20C Someone claims that the average velocity in a circular pipe in fully developed laminar flow can be determined by simply measuring the velocity at R/2 (midway between the wall surface and the centerline). Do you agree? Explain.
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8–21C Consider fully developed laminar flow in a circular pipe. If the diameter of the pipe is reduced by half while the flow rate and the pipe length are held constant, the head loss will (a) double, (b) triple, (c) quadruple, (d) increase by a factor of 8, or (e) increase by a factor of 16.
8–31 Water at 10°C (r ! 999.7 kg/m3 and m ! 1.307 . 10&3 kg/m · s) is flowing steadily in a 0.20-cm-diameter, 15-m-long pipe at an average velocity of 1.2 m/s. Determine (a) the pressure drop, (b) the head loss, and (c) the pumping power requirement to overcome this pressure drop. Answers:
8–22C What is the physical mechanism that causes the friction factor to be higher in turbulent flow?
(a) 188 kPa, (b) 19.2 m, (c) 0.71 W
8–23C
What is turbulent viscosity? What is it caused by?
8–24C The head . loss for a certain circular pipe is given by hL ! 0.0826fL(V 2/D5), where f .is the friction factor (dimensionless), L is the pipe length, V is the volumetric flow rate, and D is the pipe diameter. Determine if the 0.0826 is a dimensional or dimensionless constant. Is this equation dimensionally homogeneous as it stands?
8–32 Water at 15°C (r ! 999.1 kg/m3 and m ! 1.138 . 10&3 kg/m · s) is flowing steadily in a 30-m-long and 4-cm-diameter horizontal pipe made of stainless steel at a rate of 8 L/s. Determine (a) the pressure drop, (b) the head loss, and (c) the pumping power requirement to overcome this pressure drop.
8 L/s
8–25C Consider fully developed laminar flow in a circular pipe. If the viscosity of the fluid is reduced by half by heating while the flow rate is held constant, how will the head loss change?
30 m
8–26C How is head loss related to pressure loss? For a given fluid, explain how you would convert head loss to pressure loss. 8–27C Consider laminar flow of air in a circular pipe with perfectly smooth surfaces. Do you think the friction factor for this flow will be zero? Explain. 8–28C Explain why the friction factor is independent of the Reynolds number at very large Reynolds numbers. and m ! 0.0278 lbm/ft 8–29E Oil at 80°F (r ! 56.8 · s) is flowing steadily in a 0.5-in-diameter, 120-ft-long pipe. During the flow, the pressure at the pipe inlet and exit is measured to be 120 psi and 14 psi, respectively. Determine the flow rate of oil through the pipe assuming the pipe is (a) horizontal, (b) inclined 20° upward, and (c) inclined 20° downward. lbm/ft3
4 cm
FIGURE P8–32 8–33E Heated air at 1 atm and 100°F is to be transported in a 400-ft-long circular plastic duct at a rate of 12 ft3/s. If the head loss in the pipe is not to exceed 50 ft, determine the minimum diameter of the duct. 8–34 In fully developed laminar flow in a circular pipe, the velocity at R/2 (midway between the wall surface and the centerline) is measured to be 6 m/s. Determine the velocity at the center of the pipe. Answer: 8 m/s 8–35 The velocity profile in fully developed laminar flow in a circular pipe of inner radius R ! 2 cm, in m/s, is given by u(r) ! 4(1 & r2/R2). Determine the average and maximum velocities in the pipe and the volume flow rate.
8–30 Oil with a density of 850 kg/m3 and kinematic viscosity of 0.00062 m2/s is being discharged by a 5-mm-diameter, 40-m-long horizontal pipe from a storage tank open to the atmosphere. The height of the liquid level above the center of the pipe is 3 m. Disregarding the minor losses, determine the flow rate of oil through the pipe.
r2 u(r) = 4 a1 – –– b R2
R = 2 cm
FIGURE P8–35 8–36 Oil tank
3m 5 mm
FIGURE P8–30
Repeat Prob. 8–35 for a pipe of inner radius 7 cm.
8–37 Consider an air solar collector that is 1 m wide and 5 m long and has a constant spacing of 3 cm between the glass cover and the collector plate. Air flows at an average temperature of 45°C at a rate of 0.15 m3/s through the 1-m-wide edge of the collector along the 5-m-long passageway. Disregarding the entrance and roughness effects, determine the pressure drop in the collector. Answer: 29 Pa
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388 FLUID MECHANICS
Glass cover
Air 0.15 m3/s
atmosphere at 88 kPa. The absolute pressure 15 m before the exit is measured to be 135 kPa. Determine the flow rate of oil through the pipe if the pipe is (a) horizontal, (b) inclined 8° upward from the horizontal, and (c) inclined 8° downward from the horizontal. 135 kPa
5m Oil
15 m
1.5 cm
Collector plate
FIGURE P8–43
Insulation
FIGURE P8–37 8–38 Consider the flow of oil with r ! 894 kg/m3 and m ! 2.33 kg/m · s in a 40-cm-diameter pipeline at an average velocity of 0.5 m/s. A 300-m-long section of the pipeline passes through the icy waters of a lake. Disregarding the entrance effects, determine the pumping power required to overcome the pressure losses and to maintain the flow of oil in the pipe. 8–39 Consider laminar flow of a fluid through a square channel with smooth surfaces. Now the average velocity of the fluid is doubled. Determine the change in the head loss of the fluid. Assume the flow regime remains unchanged. 8–40 Repeat Prob. 8–39 for turbulent flow in smooth pipes for which the friction factor is given as f ! 0.184Re&0.2. What would your answer be for fully turbulent flow in a rough pipe? 8–41 Air enters a 7-m-long section of a rectangular duct of cross section 15 cm . 20 cm made of commercial steel at 1 atm and 35°C at an average velocity of 7 m/s. Disregarding the entrance effects, determine the fan power needed to overcome the pressure losses in this section of the duct. Answer: 4.9 W 7m
8–44 Glycerin at 40°C with r ! 1252 kg/m3 and m ! 0.27 kg/m · s is flowing through a 2-cm-diameter, 25-mlong pipe that discharges into the atmosphere at 100 kPa. The flow rate through the pipe is 0.035 L/s. (a) Determine the absolute pressure 25 m before the pipe exit. (b) At what angle u must the pipe be inclined downward from the horizontal for the pressure in the entire pipe to be atmospheric pressure and the flow rate to be maintained the same? 8–45 In an air heating system, heated air at 40°C and 105 kPa absolute is distributed through a 0.2 m . 0.3 m rectangular duct made of commercial steel at a rate of 0.5 m3/s. Determine the pressure drop and head loss through a 40-mlong section of the duct. Answers: 128 Pa, 93.8 m 8–46 Glycerin at 40°C with r ! 1252 kg/m3 and m ! 0.27 kg/m · s is flowing through a 5-cm-diameter horizontal smooth pipe with an average velocity of 3.5 m/s. Determine the pressure drop per 10 m of the pipe. 8–47
Reconsider Prob. 8–46. Using EES (or other) software, investigate the effect of the pipe diameter on the pressure drop for the same constant flow rate. Let the pipe diameter vary from 1 to 10 cm in increments of 1 cm. Tabulate and plot the results, and draw conclusions. 8–48E Air at 1 atm and 60°F is flowing through a 1 ft . 1 ft square duct made of commercial steel at a rate of 1200 cfm. Determine the pressure drop and head loss per ft of the duct.
15 cm Air 7 m/s
20 cm
FIGURE P8–41 8–42E Water at 60°F passes through 0.75-in-internaldiameter copper tubes at a rate of 1.2 lbm/s. Determine the pumping power per ft of pipe length required to maintain this flow at the specified rate. 8–43 Oil with r ! 876 kg/m3 and m ! 0.24 kg/m · s is flowing through a 1.5-cm-diameter pipe that discharges into the
1 ft Air 1 ft 1200 ft3/min
FIGURE P8–48E 8–49 Liquid ammonia at &20°C is flowing through a 30m-long section of a 5-mm-diameter copper tube at a rate of
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0.15 kg/s. Determine the pressure drop, the head loss, and the pumping power required to overcome the frictional losses in the tube. Answers: 4792 kPa, 743 m, 1.08 kW
into the miter elbows or to replace the sharp turns in 90° miter elbows by smooth curved bends. Which approach will result in a greater reduction in pumping power requirements?
8–50
8–58 Water is to be withdrawn from a 3-m-high water reservoir by drilling a 1.5-cm-diameter hole at the bottom surface. Disregarding the effect of the kinetic energy correction factor, determine the flow rate of water through the hole if (a) the entrance of the hole is well-rounded and (b) the entrance is sharp-edged.
Shell-and-tube heat exchangers with hundreds of tubes housed in a shell are commonly used in practice for heat transfer between two fluids. Such a heat exchanger used in an active solar hot-water system transfers heat from a water-antifreeze solution flowing through the shell and the solar collector to fresh water flowing through the tubes at an average temperature of 60°C at a rate of 15 L/s. The heat exchanger contains 80 brass tubes 1 cm in inner diameter and 1.5 m in length. Disregarding inlet, exit, and header losses, determine the pressure drop across a single tube and the pumping power required by the tube-side fluid of the heat exchanger. After operating a long time, 1-mm-thick scale builds up on the inner surfaces with an equivalent roughness of 0.4 mm. For the same pumping power input, determine the percent reduction in the flow rate of water through the tubes.
8–59 Consider flow from a water reservoir through a circular hole of diameter D at the side wall at a vertical distance H from the free surface. The flow rate through an actual hole with a sharp-edged entrance (KL ! 0.5) will be considerably less than the flow rate calculated assuming “frictionless” flow and thus zero loss for the hole. Disregarding the effect of the kinetic energy correction factor, obtain a relation for the “equivalent diameter” of the sharp-edged hole for use in frictionless flow relations.
80 tubes
Dequiv.
1.5 m
D
Frictionless flow 1 cm Water
FIGURE P8–50 Minor Losses 8–51C What is minor loss in pipe flow? How is the minor loss coefficient KL defined? 8–52C Define equivalent length for minor loss in pipe flow. How is it related to the minor loss coefficient? 8–53C The effect of rounding of a pipe inlet on the loss coefficient is (a) negligible, (b) somewhat significant, or (c) very significant. 8–54C The effect of rounding of a pipe exit on the loss coefficient is (a) negligible, (b) somewhat significant, or (c) very significant. 8–55C Which has a greater minor loss coefficient during pipe flow: gradual expansion or gradual contraction? Why? 8–56C A piping system involves sharp turns, and thus large minor head losses. One way of reducing the head loss is to replace the sharp turns by circular elbows. What is another way? 8–57C During a retrofitting project of a fluid flow system to reduce the pumping power, it is proposed to install vanes
Actual flow
FIGURE P8–59 8–60 Repeat Prob. 8–59 for a slightly rounded entrance (KL ! 0.12). 8–61 A horizontal pipe has an abrupt expansion from D1 ! 8 cm to D2 ! 16 cm. The water velocity in the smaller section is 10 m/s and the flow is turbulent. The pressure in the smaller section is P1 ! 300 kPa. Taking the kinetic energy correction factor to be 1.06 at both the inlet and the outlet, determine the downstream pressure P2, and estimate the error that would have occurred if Bernoulli’s equation had been used. Answers: 321 kPa, 28 kPa D1 = 8 cm Water
D2 = 16 cm 10 m/s 300 kPa
FIGURE P8–61 Piping Systems and Pump Selection 8–62C A piping system involves two pipes of different diameters (but of identical length, material, and roughness) connected in series. How would you compare the (a) flow rates and (b) pressure drops in these two pipes?
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8–63C A piping system involves two pipes of different diameters (but of identical length, material, and roughness) connected in parallel. How would you compare the (a) flow rates and (b) pressure drops in these two pipes? 8–64C A piping system involves two pipes of identical diameters but of different lengths connected in parallel. How would you compare the pressure drops in these two pipes? 8–65C Water is pumped from a large lower reservoir to a higher reservoir. Someone claims that if the head loss is negligible, the required pump head is equal to the elevation difference between the free surfaces of the two reservoirs. Do you agree? 8–66C A piping system equipped with a pump is operating steadily. Explain how the operating point (the flow rate and the head loss) is established. 8–67C For a piping system, define the system curve, the characteristic curve, and the operating point on a head versus flow rate chart.
8–70 A 3-m-diameter tank is initially filled with water 2 m above the center of a sharp-edged 10-cm-diameter orifice. The tank water surface is open to the atmosphere, and the orifice drains to the atmosphere. Neglecting the effect of the kinetic energy correction factor, calculate (a) the initial velocity from the tank and (b) the time required to empty the tank. Does the loss coefficient of the orifice cause a significant increase in the draining time of the tank?
Water tank
2m
3m
Sharp-edged orifice
8–68
Water at 20°C is to be pumped from a reservoir (zA ! 2 m) to another reservoir at a higher elevation (zB ! 9 m) through two 25-m-long plastic pipes connected in parallel. The diameters of the two pipes are 3 cm and 5 cm. Water is to be pumped by a 68 percent efficient motor–pump unit that draws 7 kW of electric power during operation. The minor losses and the head loss in the pipes that connect the parallel pipes to the two reservoirs are considered to be negligible. Determine the total flow rate between the reservoirs and the flow rates through each of the parallel pipes. Reservoir B zB = 9 m 25 m 3 cm Reservoir A zA = 2 m
5 cm
FIGURE P8–70 8–71 A 3-m-diameter tank is initially filled with water 2 m above the center of a sharp-edged 10-cm-diameter orifice. The tank water surface is open to the atmosphere, and the orifice drains to the atmosphere through a 100-m-long pipe. The friction coefficient of the pipe can be taken to be 0.015 and the effect of the kinetic energy correction factor can be neglected. Determine (a) the initial velocity from the tank and (b) the time required to empty the tank. 8–72 Reconsider Prob. 8–71. In order to drain the tank faster, a pump is installed near the tank exit. Determine how much pump power input is necessary to establish an average water velocity of 4 m/s when the tank is full at z ! 2 m. Also, assuming the discharge velocity to remain constant, estimate the time required to drain the tank. Someone suggests that it makes no difference whether the pump is located at the beginning or at the end of the pipe, and that the performance will be the same in either case, but
Pump
FIGURE P8–68 8–69E Water at 70°F flows by gravity from a large reservoir at a high elevation to a smaller one through a 120-ft-long, 2in-diameter cast iron piping system that includes four standard flanged elbows, a well-rounded entrance, a sharp-edged exit, and a fully open gate valve. Taking the free surface of the lower reservoir as the reference level, determine the elevation z1 of the higher reservoir for a flow rate of 10 ft3/min. Answer: 23.1 ft
Water tank 3m
FIGURE P8–72
2m Pump
4 m/s
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391 CHAPTER 8
another person argues that placing the pump near the end of the pipe may cause cavitation. The water temperature is 30°C, so the water vapor pressure is Pv ! 4.246 kPa ! 0.43 m-H2O, and the system is located at sea level. Investigate if there is the possibility of cavitation and if we should be concerned about the location of the pump. 8–73 Oil at 20°C is flowing through a vertical glass funnel that consists of a 15-cm-high cylindrical reservoir and a 1cm-diameter, 25-cm-high pipe. The funnel is always maintained full by the addition of oil from a tank. Assuming the entrance effects to be negligible, determine the flow rate of oil through the funnel and calculate the “funnel effectiveness,” which can be defined as the ratio of the actual flow rate through the funnel to the maximum flow rate for the “frictionless” case. Answers: 4.09 . 10&6 m3/s, 1.86 percent
15 cm
Oil 1 cm
25 cm
Oil
FIGURE P8–73 8–74 Repeat Prob. 8–73 assuming (a) the diameter of the pipe is doubled and (b) the length of the pipe is doubled. 8–75 Water at 15°C is drained from a large reservoir using two horizontal plastic pipes connected in series. The first pipe is 20 m long and has a 10-cm diameter, while the second pipe is 35 m long and has a 4-cm diameter. The water level in the reservoir is 18 m above the centerline of the pipe. The pipe entrance is sharp-edged, and the contraction between the two pipes is sudden. Neglecting the effect of the kinetic energy correction factor, determine the discharge rate of water from the reservoir.
8–76E A farmer is to pump water at 70°F from a river to a water storage tank nearby using a 125-ft-long, 5-in-diameter plastic pipe with three flanged 90° smooth bends. The water velocity near the river surface is 6 ft/s, and the pipe inlet is placed in the river normal to the flow direction of water to take advantage of the dynamic pressure. The elevation difference between the river and the free surface of the tank is 12 ft. For a flow rate of 1.5 ft3/s and an overall pump efficiency of 70 percent, determine the required electric power input to the pump. 8–77E
Reconsider Prob. 8–76E. Using EES (or other) software, investigate the effect of the pipe diameter on the required electric power input to the pump. Let the pipe diameter vary from 1 to 10 in, in increments of 1 in. Tabulate and plot the results, and draw conclusions. 8–78 A water tank filled with solar-heated water at 40°C is to be used for showers in a field using gravity-driven flow. The system includes 20 m of 1.5-cm-diameter galvanized iron piping with four miter bends (90°) without vanes and a wide-open globe valve. If water is to flow at a rate of 0.7 L/s through the shower head, determine how high the water level in the tank must be from the exit level of the shower. Disregard the losses at the entrance and at the shower head, and neglect the effect of the kinetic energy correction factor. 8–79 Two water reservoirs A and B are connected to each other through a 40-m-long, 2-cm-diameter cast iron pipe with a sharp-edged entrance. The pipe also involves a swing check valve and a fully open gate valve. The water level in both reservoirs is the same, but reservoir A is pressurized by compressed air while reservoir B is open to the atmosphere at 88 kPa. If the initial flow rate through the pipe is 1.2 L/s, determine the absolute air pressure on top of reservoir A. Take the water temperature to be 10°C. Answer: 733 kPa Air
40 m 2 cm
FIGURE P8–79
Water tank
18 m
20 m
FIGURE P8–75
35 m
8–80 A vented tanker is to be filled with fuel oil with r ! 920 kg/m3 and m ! 0.045 kg/m · s from an underground reservoir using a 20-m-long, 5-cm-diameter plastic hose with a slightly rounded entrance and two 90° smooth bends. The elevation difference between the oil level in the reservoir and the top of the tanker where the hose is discharged is 5 m. The capacity of the tanker is 18 m3 and the filling time is 30 min. Taking the kinetic energy correction factor at hose discharge
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to be 1.05 and assuming an overall pump efficiency of 82 percent, determine the required power input to the pump. Tanker 18 m3
20 m
5m
5 cm Pump
ligible because of the large length-to-diameter ratio and the relatively small number of components that cause minor losses. (a) Assuming the pump–motor efficiency to be 74 percent, determine the electric power consumption of the system for pumping. Would you recommend the use of a single large pump or several smaller pumps of the same total pumping power scattered along the pipeline? Explain. (b) Determine the daily cost of power consumption of the system if the unit cost of electricity is $0.06/kWh. (c) The temperature of geothermal water is estimated to drop 0.5°C during this long flow. Determine if the frictional heating during flow can make up for this drop in temperature. 8–85 Repeat Prob. 8–84 for cast iron pipes of the same diameter.
FIGURE P8–80 8–81 Two pipes of identical length and material are connected in parallel. The diameter of pipe A is twice the diameter of pipe B. Assuming the friction factor to be the same in both cases and disregarding minor losses, determine the ratio of the flow rates in the two pipes. 8–82 A certain part of cast iron piping of a water distribution system involves a parallel section. Both parallel pipes have a diameter of 30 cm, and the flow is fully turbulent. One of the branches (pipe A) is 1000 m long while the other branch (pipe B) is 3000 m long. If the flow rate through pipe A is 0.4 m3/s, determine the flow rate through pipe B. Disregard minor losses and assume the water temperature to be 15°C. Show that the flow is fully turbulent, and thus the friction factor is independent of Reynolds number. Answer:
8–86E A clothes dryer discharges air at 1 atm and 120°F at a rate of 1.2 ft3/s when its 5-in-diameter, well-rounded vent with negligible loss is not connected to any duct. Determine the flow rate when the vent is connected to a 15-ft-long, 5-indiameter duct made of galvanized iron, with three 90° flanged smooth bends. Take the friction factor of the duct to be 0.019, and assume the fan power input to remain constant.
Hot air
0.231 m3/s Clothes drier
15 ft
1000 m 0.4 m
A
3/s
5 in
30 cm
B
30 cm
FIGURE P8–86E 3000 m
FIGURE P8–82 8–83 Repeat Prob. 8–82 assuming pipe A has a halfwayclosed gate valve (KL ! 2.1) while pipe B has a fully open globe valve (KL ! 10), and the other minor losses are negligible. Assume the flow to be fully turbulent. 8–84 A geothermal district heating system involves the transport of geothermal water at 110°C from a geothermal well to a city at about the same elevation for a distance of 12 km at a rate of 1.5 m3/s in 60-cm-diameter stainless-steel pipes. The fluid pressures at the wellhead and the arrival point in the city are to be the same. The minor losses are neg-
8–87 In large buildings, hot water in a water tank is circulated through a loop so that the user doesn’t have to wait for all the water in long piping to drain before hot water starts coming out. A certain recirculating loop involves 40-m-long, 1.2-cm-diameter cast iron pipes with six 90° threaded smooth bends and two fully open gate valves. If the average flow velocity through the loop is 2.5 m/s, determine the required power input for the recirculating pump. Take the average water temperature to be 60°C and the efficiency of the pump to be 70 percent. Answer: 0.217 kW 8–88
Reconsider Prob. 8–87. Using EES (or other) software, investigate the effect of the average flow velocity on the power input to the recirculating pump.
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Let the velocity vary from 0 to 3 m/s in increments of 0.3 m/s. Tabulate and plot the results. 8–89
Repeat Prob. 8–87 for plastic pipes.
Flow Rate and Velocity Measurements
4-in-diameter pipe. A mercury manometer is used to measure the pressure difference across the orifice. If the differential height of the manometer is read to be 6 in, determine the volume flow rate of water through the pipe, the average velocity, and the head loss caused by the orifice meter.
8–90C What are the primary considerations when selecting a flowmeter to measure the flow rate of a fluid? 8–91C Explain how flow rate is measured with a Pitot-static tube, and discuss its advantages and disadvantages with respect to cost, pressure drop, reliability, and accuracy.
4 in
2 in
8–92C Explain how flow rate is measured with obstructiontype flowmeters. Compare orifice meters, flow nozzles, and Venturi meters with respect to cost, size, head loss, and accuracy.
6 in
8–93C How do positive displacement flowmeters operate? Why are they commonly used to meter gasoline, water, and natural gas? 8–94C Explain how flow rate is measured with a turbine flowmeter, and discuss how they compare to other types of flowmeters with respect to cost, head loss, and accuracy. 8–95C What is the operating principle of variable-area flowmeters (rotameters)? How do they compare to other types of flowmeters with respect to cost, head loss, and reliability? 8–96C What is the difference between the operating principles of thermal and laser Doppler anemometers? 8–97C What is the difference between laser Doppler velocimetry (LDV) and particle image velocimetry (PIV)? 8–98 The flow rate of ammonia at 10°C (r ! 624.6 kg/m3 and m ! 1.697 . 10&4 kg/m · s) through a 3-cm-diameter pipe is to be measured with a 1.5-cm-diameter flow nozzle equipped with a differential pressure gage. If the gage reads a pressure differential of 4 kPa, determine the flow rate of ammonia through the pipe, and the average flow velocity. 8–99 The flow rate of water through a 10-cm-diameter pipe is to be determined by measuring the water velocity at several locations along a cross section. For the set of measurements given in the table, determine the flow rate. r, cm
V, m/s
0 1 2 3 4 5
6.4 6.1 5.2 4.4 2.0 0.0
8–100E An orifice with a 2-in-diameter opening is used to measure the mass flow rate of water at 60°F (r ! 62.36 lbm/ft3 and m ! 7.536 . 10&4 lbm/ft · s) through a horizontal
FIGURE P8–100E 8–101E 9 in.
Repeat Prob. 8–100E for a differential height of
8–102 The flow rate of water at 20°C (r ! 998 kg/m3 and m ! 1.002 . 10&3 kg/m · s) through a 50-cm-diameter pipe is measured with an orifice meter with a 30-cm-diameter opening to be 250 L/s. Determine the pressure difference indicated by the orifice meter and the head loss. 8–103 A Venturi meter equipped with a differential pressure gage is used to measure the flow rate of water at 15°C (r ! 999.1 kg/m3) through a 5-cm-diameter horizontal pipe. The diameter of the Venturi neck is 3 cm, and the measured pressure drop is 5 kPa. Taking the discharge coefficient to be 0.98, determine the volume flow rate of water and the average velocity through the pipe. Answers: 2.35 L/s and 1.20 m/s
5 cm
3 cm
(P
Differential pressure gage
FIGURE P8–103
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394 FLUID MECHANICS
8–104
Reconsider Prob. 8–103. Letting the pressure drop vary from 1 kPa to 10 kPa, evaluate the flow rate at intervals of 1 kPa, and plot it against the pressure drop.
flow rate of liquid propane at 10°C (r ! 514.7 kg/m3) through an 8-cm-diameter vertical pipe. For a discharge coefficient of 0.98, determine the volume flow rate of propane through the pipe.
8–105 The mass flow rate of air at 20°C (r ! 1.204 kg/m3) through a 15-cm-diameter duct is measured with a Venturi meter equipped with a water manometer. The Venturi neck has a diameter of 6 cm, and the manometer has a maximum differential height of 40 cm. Taking the discharge coefficient to be 0.98, determine the maximum mass flow rate of air this Venturi meter can measure. Answer: 0.273 kg/s
8–108 A flow nozzle equipped with a differential pressure gage is used to measure the flow rate of water at 10°C (r ! 999.7 kg/m3 and m ! 1.307 . 10&3 kg/m · s) through a 3cm-diameter horizontal pipe. The nozzle exit diameter is 1.5 cm, and the measured pressure drop is 3 kPa. Determine the volume flow rate of water, the average velocity through the pipe, and the head loss.
15 cm
6 cm 3 cm
1.5 cm
Water manometer
h
(P = 3 kPa
FIGURE P8–105 8–106 Repeat Prob. 8–105 for a Venturi neck diameter of 7.5 cm. 8–107 A vertical Venturi meter equipped with a differential pressure gage shown in Fig. P8–107 is used to measure the
Differential pressure gage?
FIGURE P8–108 8–109 A 16-L kerosene tank (r ! 820 kg/m3) is filled with a 2-cm-diameter hose equipped with a 1.5-cm-diameter nozzle meter. If it takes 20 s to fill the tank, determine the pressure difference indicated by the nozzle meter. 8–110 The flow rate of water at 20°C (r ! 998 kg/m3 and m ! 1.002 . 10&3 kg/m · s) through a 4-cm-diameter pipe is measured with a 2-cm-diameter nozzle meter equipped with an inverted air–water manometer. If the manometer indicates
5 cm
(P = 7 kPa
32 cm
30 cm
4 cm
Water
8 cm
FIGURE P8–107
FIGURE P8–110
2 cm
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a differential water height of 32 cm, determine the volume flow rate of water and the head loss caused by the nozzle meter. 8–111E The volume flow rate of liquid refrigerant-134a at 10°F (r ! 83.31 lbm/ft3) is to be measured with a horizontal Venturi meter with a diameter of 5 in at the inlet and 2 in at the throat. If a differential pressure meter indicates a pressure drop of 7.4 psi, determine the flow rate of the refrigerant. Take the discharge coefficient of the Venturi meter to be 0.98.
Review Problems 8–112 The compressed air requirements of a manufacturing facility are met by a 150-hp compressor that draws in air from the outside through an 8-m-long, 20-cm-diameter duct made of thin galvanized iron sheets. The compressor takes in air at a rate of 0.27 m3/s at the outdoor conditions of 15°C and 95 kPa. Disregarding any minor losses, determine the useful power used by the compressor to overcome the frictional losses in this duct. Answer: 9.66 W Air, 0.27 m3/s 15°C, 95 kPa
20 cm
8m
Air compressor 150 hp
FIGURE P8–112 8–113 A house built on a riverside is to be cooled in summer by utilizing the cool water of the river. A 15-m-long sec-
tion of a circular stainless-steel duct of 20-cm diameter passes through the water. Air flows through the underwater section of the duct at 3 m/s at an average temperature of 15°C. For an overall fan efficiency of 62 percent, determine the fan power needed to overcome the flow resistance in this section of the duct. 8–114 The velocity profile in fully developed laminar flow in a circular pipe, in m/s, is given by u(r) ! 6(1 & 100r2), where r is the radial distance from the centerline of the pipe in m. Determine (a) the radius of the pipe, (b) the average velocity through the pipe, and (c) the maximum velocity in the pipe. 8–115E The velocity profile in a fully developed laminar flow of water at 40°F in a 80-ft-long horizontal circular pipe, in ft/s, is given by u(r) ! 0.8(1 & 625r2), where r is the radial distance from the centerline of the pipe in ft. Determine (a) the volume flow rate of water through the pipe, (b) the pressure drop across the pipe, and (c) the useful pumping power required to overcome this pressure drop. 8–116E Repeat Prob. 8–115E assuming the pipe is inclined 12° from the horizontal and the flow is uphill. 8–117 Consider flow from a reservoir through a horizontal pipe of length L and diameter D that penetrates into the side wall at a vertical distance H from the free surface. The flow rate through an actual pipe with a reentrant section (KL ! 0.8) will be considerably less than the flow rate through the hole calculated assuming “frictionless” flow and thus zero loss. Obtain a relation for the “equivalent diameter” of the reentrant pipe for use in relations for frictionless flow through a hole and determine its value for a pipe friction factor, length, and diameter of 0.018, 10 m, and 0.04 m, respectively. Assume the friction factor of the pipe to remain constant and the effect of the kinetic energy correction factor to be negligible. 8–118 Water is to be withdrawn from a 5-m-high water reservoir by drilling a well-rounded 3-cm-diameter hole with negligible loss at the bottom surface and attaching a horizontal 90° bend of negligible length. Taking the kinetic energy correction factor to be 1.05, determine the flow rate of water through the bend if (a) the bend is a flanged smooth bend
Air, 3 m/s
5m Air River
FIGURE P8–113
FIGURE P8–118
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and (b) the bend is a miter bend without vanes. Answers: (a) 0.00603 m3/s, (b) 0.00478 m3/s
8–119
In a geothermal district heating system, 10,000 kg/s of hot water must be delivered a distance of 10 km in a horizontal pipe. The minor losses are negligible, and the only significant energy loss will arise from pipe friction. The friction factor can be taken to be 0.015. Specifying a larger-diameter pipe would reduce water velocity, velocity head, pipe friction, and thus power consumption. But a larger pipe would also cost more money initially to purchase and install. Otherwise stated, there is an optimum pipe diameter that will minimize the sum of pipe cost and future electric power cost. Assume the system will run 24 h/day, every day, for 30 years. During this time the cost of electricity will remain constant at $0.06/kWh. Assume system performance stays constant over the decades (this may not be true, especially if highly mineralized water is passed through the pipeline— scale may form). The pump has an overall efficiency of 80 percent. The cost to purchase, install, and insulate a 10-km pipe depends on the diameter D and is given by Cost ! $106 D2, where D is in m. Assuming zero inflation and interest rate for simplicity and zero salvage value and zero maintenance cost, determine the optimum pipe diameter. 8–120 Water at 15°C is to be discharged from a reservoir at a rate of 18 L/s using two horizontal cast iron pipes connected in series and a pump between them. The first pipe is 20 m long and has a 6-cm diameter, while the second pipe is 35 m long and has a 4-cm diameter. The water level in the reservoir is 30 m above the centerline of the pipe. The pipe entrance is sharp-edged, and losses associated with the connection of the pump are negligible. Neglecting the effect of the kinetic energy correction factor, determine the required pumping head and the minimum pumping power to maintain the indicated flow rate.
Water tank
30 m 20 m Pump 6 cm
indicated flow rate. Let the diameter vary from 1 to 10 cm in increments of 1 cm. Tabulate and plot the results. 8–122 Two pipes of identical diameter and material are connected in parallel. The length of pipe A is twice the length of pipe B. Assuming the flow is fully turbulent in both pipes and thus the friction factor is independent of the Reynolds number and disregarding minor losses, determine the ratio of the flow rates in the two pipes. Answer: 0.707 8–123
A pipeline that transports oil at 40°C at a rate of 3 m3/s branches out into two parallel pipes made of commercial steel that reconnect downstream. Pipe A is 500 m long and has a diameter of 30 cm while pipe B is 800 m long and has a diameter of 45 cm. The minor losses are considered to be negligible. Determine the flow rate through each of the parallel pipes. A
500 m
Oil
30 cm
3 m3/s 45 cm B
800 m
FIGURE P8–123 8–124 Repeat Prob. 8–123 for hot-water flow of a district heating system at 100°C. 8–125E A water fountain is to be installed at a remote location by attaching a cast iron pipe directly to a water main through which water is flowing at 70°F and 60 psig. The entrance to the pipe is sharp-edged, and the 50-ft-long piping system involves three 90° miter bends without vanes, a fully open gate valve, and an angle valve with a loss coefficient of 5 when fully open. If the system is to provide water at a rate of 20 gal/min and the elevation difference between the pipe and the fountain is negligible, determine the minimum diameter of the piping system. Answer: 0.76 in
35 m 4 cm
50 ft
FIGURE P8–120 8–121
Reconsider Prob. 8–120. Using EES (or other) software, investigate the effect of the second pipe diameter on the required pumping head to maintain the
Water main
60 psig
FIGURE P8–125E
20 gpm
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397 CHAPTER 8
8–126E
Repeat Prob. 8–125E for plastic pipes.
8–127 In a hydroelectric power plant, water at 20°C is supplied to the turbine at a rate of 0.8 m3/s through a 200-mlong, 0.35-m-diameter cast iron pipe. The elevation difference between the free surface of the reservoir and the turbine discharge is 70 m, and the combined turbine–generator efficiency is 84 percent. Disregarding the minor losses because of the large length-to-diameter ratio, determine the electric power output of this plant. 8–128 In Prob. 8–127, the pipe diameter is tripled in order to reduce the pipe losses. Determine the percent increase in the net power output as a result of this modification. 8–129E The drinking water needs of an office are met by large water bottles. One end of a 0.35-in-diameter, 6-ft-long plastic hose is inserted into the bottle placed on a high stand, while the other end with an on/off valve is maintained 3 ft below the bottom of the bottle. If the water level in the bottle is 1 ft when it is full, determine how long it will take to fill an 8-oz glass (! 0.00835 ft3) (a) when the bottle is first opened and (b) when the bottle is almost empty. Take the total minor loss coefficient, including the on/off valve, to be 2.8 when it is fully open. Assume the water temperature to be the same as the room temperature of 70°F. Answers: (a) 2.4 s, (b) 2.8 s 6 ft
0.35 in
1 ft
important. So he used the entire 12-ft-long tube. Assuming the turns or constrictions in the tube are not significant (being very optimistic) and the same elevation is maintained, determine the time it takes to fill a glass of water for both cases. 8–132 A circular water pipe has an abrupt expansion from diameter D1 ! 15 cm to D2 ! 20 cm. The pressure and the average water velocity in the smaller pipe are P1 ! 120 kPa and 10 m/s, respectively, and the flow is turbulent. By applying the continuity, momentum, and energy equations and disregarding the effects of the kinetic energy and momentumflux correction factors, show that the loss coefficient for sudden expansion is KL ! (1 & D12/D22)2, and calculate KL and P2 for the given case.
V1 = 10 m/s
D1
D2
FIGURE P8–132 8–133 The water at 20°C in a 10-m-diameter, 2-m-high aboveground swimming pool is to be emptied by unplugging a 3-cm-diameter, 25-m-long horizontal plastic pipe attached to the bottom of the pool. Determine the initial rate of discharge of water through the pipe and the time it will take to empty the swimming pool completely assuming the entrance to the pipe is well-rounded with negligible loss. Take the friction factor of the pipe to be 0.022. Using the initial discharge velocity, check if this is a reasonable value for the friction factor. Answers: 1.01 L/s, 86.7 h 10 m
3 ft 2m
Swimming pool
FIGURE P8–129E
25 m
FIGURE P8–133
8–130E
Reconsider Prob. 8–129E. Using EES (or other) software, investigate the effect of the hose diameter on the time required to fill a glass when the bottle is full. Let the diameter vary from 0.2 to 2 in, in increments of 0.2 in. Tabulate and plot the results. 8–131E Reconsider Prob. 8–129E. The office worker who set up the siphoning system purchased a 12-ft-long reel of the plastic tube and wanted to use the whole thing to avoid cutting it in pieces, thinking that it is the elevation difference that makes siphoning work, and the length of the tube is not
3 cm
8–134
Reconsider Prob. 8–133. Using EES (or other) software, investigate the effect of the discharge pipe diameter on the time required to empty the pool completely. Let the diameter vary from 1 to 10 cm, in increments of 1 cm. Tabulate and plot the results. 8–135 Repeat Prob. 8–133 for a sharp-edged entrance to the pipe with KL ! 0.5. Is this “minor loss” truly “minor” or not?
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398 FLUID MECHANICS
8–136 A system that consists of two interconnected cylindrical tanks with D1 ! 30 cm and D2 ! 12 cm is to be used to determine the discharge coefficient of a short D0 ! 5 mm diameter orifice. At the beginning (t ! 0 s), the fluid heights in the tanks are h1 ! 50 cm and h2 ! 15 cm, as shown in Fig. P8–136. If it takes 170 s for the fluid levels in the two tanks to equalize and the flow to stop, determine the discharge coefficient of the orifice. Disregard any other losses associated with this flow.
garding entrance effects and velocity heads, obtain a relation for the variation of fluid depth in the tank with time. 8–138 A student is to determine the kinematic viscosity of an oil using the system shown in Prob. 8–137. The initial fluid height in the tank is H ! 40 cm, the tube diameter is d ! 6 mm, the tube length is L ! 0.65 m, and the tank diameter is D ! 0.63 m. The student observes that it takes 2842 s for the fluid level in the tank to drop to 36 cm. Find the fluid viscosity.
Design and Essay Problems 8–139 Electronic boxes such as computers are commonly cooled by a fan. Write an essay on forced air cooling of electronic boxes and on the selection of the fan for electronic devices. h1
h
8–140 Design an experiment to measure the viscosity of liquids using a vertical funnel with a cylindrical reservoir of height h and a narrow flow section of diameter D and length L. Making appropriate assumptions, obtain a relation for viscosity in terms of easily measurable quantities such as density and volume flow rate. Is there a need for the use of a correction factor?
h2
Tank 1
Tank 2 Orifice
FIGURE P8–136 8–137 A highly viscous liquid discharges from a large container through a small-diameter tube in laminar flow. Disre-
H
D
FIGURE P8–137
Discharge tube
L
d
8–141 A pump is to be selected for a waterfall in a garden. The water collects in a pond at the bottom, and the elevation difference between the free surface of the pond and the location where the water is discharged is 3 m. The flow rate of water is to be at least 8 L/s. Select an appropriate motor– pump unit for this job and identify three manufacturers with product model numbers and prices. Make a selection and explain why you selected that particular product. Also estimate the cost of annual power consumption of this unit assuming continuous operation. 8–142 During a camping trip you notice that water is discharged from a high reservoir to a stream in the valley through a 30-cm-diameter plastic pipe. The elevation difference between the free surface of the reservoir and the stream is 70 m. You conceive the idea of generating power from this water. Design a power plant that will produce the most power from this resource. Also, investigate the effect of power generation on the discharge rate of water. What discharge rate will maximize the power production?
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CHAPTER
9
D I F F E R E N T I A L A N A LY S I S OF FLUID FLOW
I
n this chapter we derive the differential equations of fluid motion, namely, conservation of mass (the continuity equation) and Newton’s second law (the Navier–Stokes equation). These equations apply to every point in the flow field and thus enable us to solve for all details of the flow everywhere in the flow domain. Unfortunately, most differential equations encountered in fluid mechanics are very difficult to solve and often require the aid of a computer. Also, these equations must be combined when necessary with additional equations, such as an equation of state and an equation for energy and/or species transport. We provide a step-by-step procedure for solving this set of differential equations of fluid motion and obtain analytical solutions for several simple examples. We also introduce the concept of the stream function; curves of constant stream function turn out to be streamlines in two-dimensional flow fields.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
Understand how the differential equations of mass and momentum conservation are derived Calculate the stream function and pressure field, and plot streamlines for a known velocity field Obtain analytical solutions of the equations of motion for simple flow fields
399
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400 FLUID MECHANICS
9–1
Control volume Flow out
Flow in Flow out
→
F
(a) Flow domain Flow out
Flow in Flow out
→
F (b)
FIGURE 9–1 (a) In control volume analysis, the interior of the control volume is treated like a black box, but (b) in differential analysis, all the details of the flow are solved at every point within the flow domain.
■
INTRODUCTION
In Chap. 5, we derived control volume versions of the laws of conservation of mass and energy, and in Chap. 6 we did the same for momentum. The control volume technique is useful when we are interested in the overall features of a flow, such as mass flow rate into and out of the control volume or net forces applied to bodies. An example is sketched in Fig. 9–1a for the case of wind flowing around a satellite dish. A rectangular control volume is taken around the vicinity of the satellite dish, as sketched. If we know the air velocity along the entire control surface, we can calculate the net reaction force on the satellite dish without ever knowing any details about its geometry. The interior of the control volume is in fact treated like a “black box” in control volume analysis—we cannot obtain detailed knowledge about flow properties such as velocity or pressure at points inside the control volume. Differential analysis, on the other hand, involves application of differential equations of fluid motion to any and every point in the flow field over a region called the flow domain. You can think of the differential technique as the analysis of millions of tiny control volumes stacked end to end and on top of each other all throughout the flow field. In the limit as the number of tiny control volumes goes to infinity, and the size of each control volume shrinks to a point, the conservation equations simplify to a set of partial differential equations that are valid at any point in the flow. When solved, these differential equations yield details about the velocity, density, pressure, etc., at every point throughout the entire flow domain. In Fig. 9–1b, for example, differential analysis of airflow around the satellite dish yields streamline shapes, a detailed pressure distribution around the dish, etc. From these details, we can integrate to find gross features of the flow such as the net force on the satellite dish. In a fluid flow problem such as the one illustrated in Fig. 9–1 in which air density and temperature changes are insignificant, it is sufficient to solve two differential equations of motion—conservation of mass and Newton’s second law (conservation of linear momentum). For three-dimensional incompressible flow, there are four unknowns (velocity components u, v, w, and pressure P) and four equations (one from conservation of mass, which is a scalar equation, and three from Newton’s second law, which is a vector equation). As we shall see, the equations are coupled, meaning that some of the variables appear in all four equations; the set of differential equations must therefore be solved simultaneously for all four unknowns. In addition, boundary conditions for the variables must be specified at all boundaries of the flow domain, including inlets, outlets, and walls. Finally, if the flow is unsteady, we must march our solution along in time as the flow field changes. You can see how differential analysis of fluid flow can become quite complicated and difficult. Computers are a tremendous help here, as discussed in Chap. 15. Nevertheless, there is much we can do analytically, and we start by deriving the differential equation for conservation of mass.
9–2
■
CONSERVATION OF MASS— THE CONTINUITY EQUATION
Through application of the Reynolds transport theorem (Chap. 4), we have the following general expression for conservation of mass as applied to a control volume:
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401 CHAPTER 9
Conservation of mass for a CV: 0"
x1
!
#r dV $ #t CV
!
→
→
rV % n dA
z1 (9–1)
CS
y1 CV
Recall that Eq. 9–1 is valid for both fixed and moving control volumes, provided that the velocity vector is the absolute velocity (as seen by a fixed observer). When there are well-defined inlets and outlets, Eq. 9–1 can be rewritten as #r # # dV " a m & a m #t in out CV
!
(9–2)
In words, the net rate of change of mass within the control volume is equal to the rate at which mass flows into the control volume minus the rate at which mass flows out of the control volume. Equation 9–2 applies to any control volume, regardless of its size. To generate a differential equation for conservation of mass, we imagine the control volume shrinking to infinitesimal size, with dimensions dx, dy, and dz (Fig. 9–2). In the limit, the entire control volume shrinks to a point in the flow.
Derivation Using the Divergence Theorem
The quickest and most straightforward way to derive the differential form of conservation of mass is to apply the divergence theorem to Eq. 9–1. The divergence theorem is also called Gauss’s theorem, named after the German mathematician Johann Carl Friedrich Gauss (1777–1855). The divergence theorem allows us to transform a volume integral of the divergence of a vector into→ an area integral over→the surface that→ defines the volume. For → any vector G , the divergence of G is defined as ! · G , and the divergence theorem can be written as
! § % G dV " " G % n dA →
Divergence theorem:
→
→
V
→
(9–3)
A
The circle on the area integral is used to emphasize that the integral must be evaluated around the entire closed area A that surrounds volume V. Note that the control surface of Eq. 9–1 is a closed area, even though we do not always add the circle to the integral symbol. Equation 9–3 applies to →any vol→ ume, so we choose the control volume of Eq. 9–1. We also let G " rV → since G can be any vector. Substitution of Eq. 9–3 into Eq. 9–1 converts the area integral into a volume integral,
!
#r dV $ #t CV
0"
!
CV
→
→
§ % arVb dV
We now combine the two volume integrals into one,
!
CV
→ #r → $ § % arVbd dV " 0 #t
c
(9–4)
Finally, we argue that Eq. 9–4 must hold for any control volume regardless of its size or shape. This is possible only if the integrand (the terms within
y x z
dy dz dx
FIGURE 9–2 To derive a differential conservation equation, we imagine shrinking a control volume to infinitesimal size.
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402 FLUID MECHANICS
square brackets) is identically zero. Hence, we have a general differential equation for conservation of mass, better known as the continuity equation: Continuity equation:
→ #r → $ § % arVb " 0 #t
(9–5)
Equation 9–5 is the compressible form of the continuity equation since we have not assumed incompressible flow. It is valid at any point in the flow domain.
Derivation Using an Infinitesimal Control Volume
v r w
dy
u P dz
y
dx x
z
FIGURE 9–3 A small box-shaped control volume centered at point P is used for derivation of the differential equation for conservation of mass in Cartesian coordinates; the blue dots indicate the center of each face.
We derive the continuity equation in a different way, by starting with a control volume on which we apply conservation of mass. Consider an infinitesimal box-shaped control volume aligned with the axes in Cartesian coordinates (Fig. 9–3). The dimensions of the box are dx, dy, and dz, and the center of the box is shown at some arbitrary point P from the origin (the box can be located anywhere in the flow field). At the center of the box we define the density as r and the velocity components as u, v, and w, as shown. At locations away from the center of the box, we use a Taylor series expansion about the center of the box (point P). [The series expansion is named in honor of its creator, the English mathematician Brook Taylor (1685–1731).] For example, the center of the right-most face of the box is located a distance dx/2 from the middle of the box in the x-direction; the value of ru at that point is (ru)center of right face " ru $
#(ru) dx 1 #2(ru) dx 2 a b $ p $ #x 2 2! #x 2 2
(9–6)
As the box representing the control volume shrinks to a point, however, second-order and higher terms become negligible. For example, suppose dx/L " 10&3, where L is some characteristic length scale of the flow domain. Then (dx/L)2 " 10&6, a factor of a thousand less than dx/L. In fact, the smaller dx, the better the assumption that second-order terms are negligible. Applying this truncated Taylor series expansion to the density times the normal velocity component at the center point of each of the six faces of the box, we have (ru)center of right face # ru $
#(ru) dx #x 2
(ru)center of left face # ru &
#( ru) dx #x 2
Center of front face:
(rw)center of front face # rw $
#(rw) dz #z 2
Center of rear face:
( rw)center of rear face # rw &
#(rw) dz #z 2
Center of top face:
( rv)center of top face # rv $
#(rv) dy #y 2
( rv)center of bottom face # rv &
#(rv) dy #y 2
Center of right face: Center of left face:
Center of bottom face:
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403 CHAPTER 9
The mass flow rate into or out of one of the faces is equal to the density times the normal velocity component at the center point of the face times . the surface area of the face. In other words, m " rVn A at each face, where Vn is the magnitude of the normal velocity through the face and A is the surface area of the face (Fig. 9–4). The mass flow rate through each face of our infinitesimal control volume is illustrated in Fig. 9–5. We could construct truncated Taylor series expansions at the center of each face for the remaining (nonnormal) velocity components as well, but this is unnecessary since these components are tangential to the face under consideration. For example, the value of rv at the center of the right face can be estimated by a similar expansion, but since v is tangential to the right face of the box, it contributes nothing to the mass flow rate into or out of that face. As the control volume shrinks to a point, the value of the volume integral on the left-hand side of Eq. 9–2 becomes
A = surface area Vn = average normal velocity component y x z
FIGURE 9–4 The mass flow rate through a surface is equal to rVnA.
Rate of change of mass within CV:
!
#r #r dV # dx dy dz #t #t CV
(9–7)
since the volume of the box is dx dy dz. We now apply the approximations of Fig. 9–5 to the right-hand side of Eq. 9–2. We add up all the mass flow rates into and out of the control volume through the faces. The left, bottom, and back faces contribute to mass inflow, and the first term on the righthand side of Eq. 9–2 becomes Net mass flow rate into CV: #(ru) dx #(rv) dy #(rw) dz # a m # ¢ru & #x 2 ≤ dy dz $ ¢rv & #y 2 ≤ dx dz $ ¢rw & #z 2 ≤ dx dy in
bottom face
y
left face
rear face
arv + ∂(rv) dyb dx dz ∂y
x z
2
arw – ∂(rw) dzb dx dy ∂z
dy
aru – ∂(ru) dxb dy dz ∂x
2
arw + ∂(rw) dzb dx dy ∂z
2
2
aru + ∂(ru) dxb dy dz ∂x
dz dx
arv – ∂(rv) dyb dx dz ∂y
2
2
FIGURE 9–5 The inflow or outflow of mass through each face of the differential control volume; the blue dots indicate the center of each face.
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404 FLUID MECHANICS
The Divergence Operation Cartesian coordinates: → → ∂ ∂ ∂ • (rV ) = (ru)) + (rv)) + (rw) ∂x ∂y ∂z ∆
(rV ) = 1 ∂(rrur)
∆
r
∂r
+
1 ∂(ruu) ∂(ruz) + ∂z r ∂u
out
right face
top face
•
→
#(ru) dx #(rv) dy #(rw) dz # a m # ¢ru $ #x 2 ≤ dy dz $ ¢rv $ #y 2 ≤ dx dz $ ¢rw $ #z 2 ≤ dx dy
→
Net mass flow rate out of CV:
Cylindrical coordinates:
Similarly, the right, top, and front faces contribute to mass outflow, and the second term on the right-hand side of Eq. 9–2 becomes
front face
We substitute Eq. 9–7 and these two equations for mass flow rate into Eq. 9–2. Many of the terms cancel each other out; after combining and simplifying the remaining terms, we are left with #(ru) #r #( rv) #(rw) dx dy dz " & dx dy dz & dx dy dz & dx dy dz #t #x #y #z
FIGURE 9–6 The divergence operation in Cartesian and cylindrical coordinates.
The volume of the box, dx dy dz, appears in each term and can be eliminated. After rearrangement we end up with the following differential equation for conservation of mass in Cartesian coordinates: Continuity equation in Cartesian coordinates: #r #(ru) #( rv) #(rw) $ $ $ "0 #t #x #y #z
(9–8)
Equation 9–8 is the compressible form of the continuity equation in Cartesian coordinates. It can be written in more compact form by recognizing the divergence operation (Fig. 9–6), yielding exactly the same equation as Eq. 9–5. EXAMPLE 9–1 Cylinder y
L(t)
v r(t)
L bottom
Piston Time t Time t = 0
VP
FIGURE 9–7 Fuel and air being compressed by a piston in a cylinder of an internal combustion engine.
Compression of an Air–Fuel Mixture
An air–fuel mixture is compressed by a piston in a cylinder of an internal combustion engine (Fig. 9–7). The origin of coordinate y is at the top of the cylinder, and y points straight down as shown. The piston is assumed to move up at constant speed VP. The distance L between the top of the cylinder and the piston decreases with time according to the linear approximation L " Lbottom & VPt, where Lbottom is the location of the piston when it is at the bottom of its cycle at time t " 0, as sketched in Fig. 9–7. At t " 0, the density of the air–fuel mixture in the cylinder is everywhere equal to r(0). Estimate the density of the air–fuel mixture as a function of time and the given parameters during the piston’s up stroke.
SOLUTION The density of the air–fuel mixture is to be estimated as a function of time and the given parameters in the problem statement. Assumptions 1 Density varies with time, but not space; in other words, the density is uniform throughout the cylinder at any given time, but changes with time: r " r(t). 2 Velocity component v varies with y and t, but not with x or z; in other words v " v(y, t ) only. 3 u " w " 0. 4 No mass escapes from the cylinder during the compression. Analysis First we need to establish an expression for velocity component v as a function of y and t. Clearly v " 0 at y " 0 (the top of the cylinder), and v "&VP at y " L. For simplicity, we assume that v varies linearly between these two boundary conditions, Vertical velocity component:
v " &VP
y L
(1)
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405 CHAPTER 9
where L is a function of time, as given. The compressible continuity equation in Cartesian coordinates (Eq. 9–8) is appropriate for solution of this problem.
#r #(ru) #(rv) #(rw) $ $ $ "0 #t #x #y #z
0 since u " 0
0 since w " 0
→
#r #(rv) $ "0 #t #y
By assumption 1, however, density is not a function of y and can therefore come out of the y-derivative. Substituting Eq. 1 for v and the given expression for L, differentiating, and simplifying, we obtain
#r y VP VP #v # " &r " &r ¢&VP ≤ " r " r #t #y #y L L L bottom & VPt
(2)
By assumption 1 again, we replace #r/#t by dr/dt in Eq. 2. After separating variables we obtain an expression that can be integrated analytically, r
dr " r r"r(0)
!
!
t
t"0
VP dt L bottom & VPt
→
ln
r L bottom " ln r(0) L bottom & VPt
5 (3)
Finally then, we have the desired expression for r as a function of time,
r " r(0)
L bottom L bottom & VPt
4 (4)
In keeping with the convention of nondimensionalizing results, Eq. 4 can be rewritten as
r 1 " r(0) 1 & VPt/L bottom
→
R* !
1 1 " t*
(5)
where r* " r/r(0) and t* " VPt/Lbottom. Equation 5 is plotted in Fig. 9–8. Discussion At t* " 1, the piston hits the top of the cylinder and r goes to infinity. In an actual internal combustion engine, the piston stops before reaching the top of the cylinder, forming what is called the clearance volume, which typically constitutes 4 to 12 percent of the maximum cylinder volume. The assumption of uniform density within the cylinder is the weakest link in this simplified analysis. In reality, r may be a function of both space and time.
Alternative Form of the Continuity Equation
We expand Eq. 9–5 by using the product rule on the divergence term, (9–9)
→ → → #r → #r → → $ § % (rV) " $ V % §r $ r§ % V " 0 #t #t Material derivative of r
Recognizing the material derivative in Eq. 9–9 (see Chap. 4), and dividing by r, we write the compressible continuity equation in an alternative form, Alternative form of the continuity equation: 1 Dr → → $ §%V"0 r Dt
(9–10)
Equation 9–10 shows that as we follow a fluid element through the flow → → field (we call this a material element), its density changes as § · V changes
r* 3
2
1 0
0.2
0.4
0.6
0.8
1
t*
FIGURE 9–8 Nondimensional density as a function of nondimensional time for Example 9–1.
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406 FLUID MECHANICS
(Fig. 9–9). On the other hand, if changes in the density of the material element are negligibly small compared to the magnitudes of the velocity gradients → → in § · V as the element moves around, r&1Dr/Dt $ 0, and the flow is approximated as incompressible.
Continuity Equation in Cylindrical Coordinates
Many problems in fluid mechanics are more conveniently solved in cylindrical coordinates (r, u, z) (often called cylindrical polar coordinates), rather than in Cartesian coordinates. For simplicity, we introduce cylindrical coordinates in two dimensions first (Fig. 9–10a). By convention, r is the radial distance from the origin to some point (P), and u is the angle measured from the x-axis (u is always defined as mathematically positive in the counterclockwise direction). Velocity components, ur and uu, and unit vec→ → tors, e r and e u, are also shown in Fig. 9–10a. In three dimensions, imagine sliding everything in Fig. 9–10a out of the page along the z-axis (normal to the xy-plane) by some distance z. We have attempted to draw this in Fig. 9–10b. In three dimensions, we have a third velocity component, uz, and a → third unit vector, e z, also sketched in Fig. 9–10b. The following coordinate transformations are obtained from Fig. 9–10:
Streamline
FIGURE 9–9 As a material element moves through a flow field, its density changes according to Eq. 9–10.
Coordinate transformations: y
r " 2x 2 $ y 2
uu ur
→
eu →
r
P
er
y u
x
x
(a) y
x " r cos u
y " r sin u
u " tan &1
y x
(9–11)
Coordinate z is the same in cylindrical and Cartesian coordinates. To obtain an expression for the continuity equation in cylindrical coordinates, we have two choices. First, we can use Eq. 9–5 directly, since it was derived without regard to our choice of coordinate system. We simply look up the expression for the divergence operator in cylindrical coordinates in a vector calculus book (e.g., Spiegel, 1968; see also Fig. 9–6). Second, we can draw a three-dimensional infinitesimal fluid element in cylindrical coordinates and analyze mass flow rates into and out of the element, similar to what we did before in Cartesian coordinates. Either way, we end up with Continuity equation in cylindrical coordinates:
uu ur
→
eu
#r 1 #(rru r) 1 #(ru u) #(ru z) $ $ $ "0 r #r r #u #t #z
(9–12)
Details of the second method can be found in Fox and McDonald (1998). r →
z
u uz
ez
→
P
er x
z
(b)
FIGURE 9–10 Velocity components and unit vectors in cylindrical coordinates: (a) twodimensional flow in the xy- or ruplane, (b) three-dimensional flow.
Special Cases of the Continuity Equation
We now look at two special cases, or simplifications, of the continuity equation. In particular, we first consider steady compressible flow, and then incompressible flow.
Special Case 1: Steady Compressible Flow
If the flow is compressible but steady, #/#t of any variable is equal to zero. Thus, Eq. 9–5 reduces to Steady continuity equation:
→
→
§ % (rV) " 0
(9–13)
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407 CHAPTER 9
In Cartesian coordinates, Eq. 9–13 reduces to #(ru) #(rv) #(rw) $ $ "0 #x #y #z
(9–14)
In cylindrical coordinates, Eq. 9–13 reduces to 1 #(rru r) 1 #(ru u) #(ru z) $ $ "0 r #r r #u #z
(9–15)
Special Case 2: Incompressible Flow
If the flow is approximated as incompressible, density is not a function of time or space. Thus #r/#t $ 0 in Eq. 9–5, and r can be taken outside of the divergence operator. Equation 9–5 therefore reduces to Incompressible continuity equation:
→
→
§%V"0
(9–16)
The same result is obtained if we start with Eq. 9–10 and recognize that for an incompressible flow, density does not change appreciably following a fluid particle, as pointed out previously. Thus the material derivative of r is approximately zero, and Eq. 9–10 reduces immediately to Eq. 9–16. You may have noticed that no time derivatives remain in Eq. 9–16. We conclude from this that even if the flow is unsteady, Eq. 9–16 applies at any instant in time. Physically, this means that as the velocity field changes in one part of an incompressible flow field, the entire rest of the flow field immediately adjusts to the change such that Eq. 9–16 is satisfied at all times. For compressible flow this is not the case. In fact, a disturbance in one part of the flow is not even felt by fluid particles some distance away until the sound wave from the disturbance reaches that distance. Very loud noises, such as that from a gun or explosion, generate a shock wave that actually travels faster than the speed of sound. (The shock wave produced by an explosion is illustrated in Fig. 9–11.) Shock waves and other manifestations of compressible flow are discussed in Chap. 12. In Cartesian coordinates, Eq. 9–16 is
Observer
Pow!
Shock wave
Incompressible continuity equation in Cartesian coordinates: #u #v #w $ $ "0 #x #y #z
(9–17)
Equation 9–17 is the form of the continuity equation you will probably encounter most often. It applies to steady or unsteady, incompressible, three-dimensional flow, and you would do well to memorize it. In cylindrical coordinates, Eq. 9–16 is Incompressible continuity equation in cylindrical coordinates: 1 #(ru r) 1 #(u u) #(u z) $ $ "0 r #r r #u #z
EXAMPLE 9–2
(9–18)
Design of a Compressible Converging Duct
A two-dimensional converging duct is being designed for a high-speed wind tunnel. The bottom wall of the duct is to be flat and horizontal, and the top wall is to be curved in such a way that the axial wind speed u increases
FIGURE 9–11 The disturbance from an explosion is not felt until the shock wave reaches the observer.
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408 FLUID MECHANICS ∆x = 2.0 m
approximately linearly from u1 " 100 m/s at section (1) to u2 " 300 m/s at section (2) (Fig. 9–12). Meanwhile, the air density r is to decrease approximately linearly from r1 " 1.2 kg/m3 at section (1) to r2 " 0.85 kg/m3 at section (2). The converging duct is 2.0 m long and is 2.0 m high at section (1). (a) Predict the y-component of velocity, v(x, y), in the duct. (b) Plot the approximate shape of the duct, ignoring friction on the walls. (c) How high should the duct be at section (2), the exit of the duct?
2.0 m
y (1)
x
(2)
FIGURE 9–12 Converging duct, designed for a highspeed wind tunnel (not to scale).
SOLUTION For given velocity component u and density r, we are to predict velocity component v, plot an approximate shape of the duct, and predict its height at the duct exit. Assumptions 1 The flow is steady and two-dimensional in the xy-plane. 2 Friction on the walls is ignored. 3 Axial velocity u increases linearly with x, and density r decreases linearly with x. Properties The fluid is air at room temperature (25°C). The speed of sound is about 346 m/s, so the flow is subsonic, but compressible. Analysis (a) We write expressions for u and r, forcing them to be linear in x, where
u " u1 $ Cu x
Cu "
u 2 & u 1 (300 & 100) m/s " " 100 s &1 (1) 'x 2.0 m
and
where
r " r1 $ Cr x
Cr "
r 2 & r 1 (0.85 & 1.2) kg/m3 " 'x 2.0 m
(2)
" &0.175 kg/m4 The steady continuity equation (Eq. 9–14) for this two-dimensional compressible flow simplifies to
→
#(ru) #(rv) #(rw) $ $ "0 #x #y #z
#(rv) #(ru) "& #y #x
(3)
0 (2-D)
Substituting Eqs. 1 and 2 into Eq. 3 and noting that Cu and Cr are constants,
#[(r 1 $ C r x)(u 1 $ C u x)] #(rv) "& " &(r 1C u $ u 1C r) & 2C uC r x #y #x
2
Integration with respect to y gives
1.5 y
Top wall
rv " &(r 1C u $ u 1C r)y & 2C uC r xy $ f (x)
(4)
Note that since the integration is a partial integration, we have added an arbitrary function of x instead of simply a constant of integration. Next, we apply boundary conditions. We argue that since the bottom wall is flat and horizontal, v must equal zero at y " 0 for any x. This is possible only if f (x) " 0. Solving Eq. 4 for v gives
1
0.5
v"
0 0
0.5
1 x
1.5
2
Bottom wall
FIGURE 9–13 Streamlines for the converging duct of Example 9–2.
&(r 1C u $ u 1C r)y & 2C uC r xy r
→
v!
"(R1Cu # u1CR)y "2CuCR xy R1 # CR x
(5)
(b) Using Eqs. 1 and 5 and the technique described in Chap. 4, we plot several streamlines between x " 0 and x " 2.0 m in Fig. 9–13. The streamline starting at x " 0, y " 2.0 m approximates the top wall of the duct.
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409 CHAPTER 9
(c) At section (2), the top streamline crosses y ! 0.941 m at x ! 2.0 m. Thus, the predicted height of the duct at section (2) is 0.941 m. Discussion You can verify that the combination of Eqs. 1, 2, and 5 satisfies the continuity equation. However, this alone does not guarantee that the density and velocity components will actually follow these equations if the duct were to be built as designed here. The actual flow depends on the pressure drop between sections (1) and (2); only one unique pressure drop can yield the desired flow acceleration. Temperature may also change considerably in this kind of compressible flow in which the air accelerates toward sonic speeds.
EXAMPLE 9–3
Incompressibility of an Unsteady Two-Dimensional Flow
Consider the velocity field of Example 4–5—an →unsteady, two-dimensional → velocity field given by V ! (u, v) ! (0.5 " 0.8x)i " [1.5 " 2.5 sin (vt) # → 0.8y]j , where angular frequency v is equal to 2p rad/s (a physical frequency of 1 Hz). Verify that this flow field can be approximated as incompressible.
SOLUTION We are to verify that a given velocity field is incompressible. Assumptions 1 The flow is two-dimensional, implying no z-component of velocity and no variation of u or v with z. Analysis The components of velocity in the x- and y-directions, respectively, are u ! 0.5 " 0.8x
and
v ! 1.5 " 2.5 sin (vt) # 0.8y
If the flow is incompressible, Eq. 9–16 must apply. More specifically, in Cartesian coordinates Eq. 9–17 must apply. Let’s check:
$u $v $w " " !0 $x $y $z
→
0.8 # 0.8 ! 0
F
F
F 0.8
#0.8 0 since 2-D
So we see that the incompressible continuity equation is indeed satisfied at any instant in time, and this flow field may be approximated as incompressible. Discussion Although there is an unsteady term in v, it has no y-derivative and drops out of the continuity equation.
EXAMPLE 9–4
Finding a Missing Velocity Component
Two velocity components of a steady, incompressible, three-dimensional flow field are known, namely, u ! ax 2 " by 2 " cz 2 and w ! axz " byz 2, where a, b, and c are constants. The y velocity component is missing (Fig. 9–14). Generate an expression for v as a function of x, y, and z.
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Missing: y velocity component If found, call 1-800-CON-UITY
SOLUTION We are to find the y-component of velocity, v, using given expressions for u and w. Assumptions 1 The flow is steady. 2 The flow is incompressible. Analysis Since the flow is steady and incompressible, and since we are working in Cartesian coordinates, we apply Eq. 9–17 to the flow field,
FIGURE 9–14 The continuity equation can be used to find a missing velocity component.
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410 FLUID MECHANICS
Condition for incompressibility: #v #u #w "& & #y #x #z
→
#v " &3ax & 2byz #y
F
F 2ax
ax $ 2byz
Next we integrate with respect to y. Since the integration is a partial integration, we add some arbitrary function of x and z instead of a simple constant of integration.
Solution:
v ! "3axy " by2 z # f(x, z)
Discussion Any function of x and z yields a v that satisfies the incompressible continuity equation, since there are no derivatives of v with respect to x or z in the continuity equation.
EXAMPLE 9–5
Two-Dimensional, Incompressible, Vortical Flow
Consider a two-dimensional, incompressible flow in cylindrical coordinates; the tangential velocity component is uu " K/r, where K is a constant. This represents a class of vortical flows. Generate an expression for the other velocity component, ur .
uu uu =
K r
r
SOLUTION For a given tangential velocity component, we are to generate an expression for the radial velocity component. Assumptions 1 The flow is two-dimensional in the xy- (ru-) plane (velocity is not a function of z, and uz " 0 everywhere). 2 The flow is incompressible. Analysis The incompressible continuity equation (Eq. 9–18) for this twodimensional case simplifies to 1 #(ru r) 1 #u u #u z $ $ "0 r #r r #u #z
ur = 0
→
V
#u u #(ru r) "& #r #u
(1)
0 (2-D)
The given expression for uu is not a function of u, and therefore Eq. 1 reduces to
(a) uu uu =
K r
r
ur =
C r (b)
FIGURE 9–15 Streamlines and velocity profiles for (a) a line vortex flow and (b) a spiraling line vortex/sink flow.
#(ru r) "0 #r
→
ru r " f (u, t)
(2)
where we have introduced an arbitrary function of u and t instead of a constant of integration, since we performed a partial integration with respect to r. Solving for ur ,
ur "
f(u, t) r
(3)
Thus, any radial velocity component of the form given by Eq. 3 yields a twodimensional, incompressible velocity field that satisfies the continuity equation. We discuss some specific cases. The simplest case is when f(u, t) " 0 (ur " 0, uu " K/r). This yields the line vortex discussed in Chap. 4, as sketched in Fig. 9–15a. Another simple case is when f(u, t) " C, where C is a constant. This yields a radial velocity whose magnitude decays as 1/r. For negative C, imagine a spiraling line vortex/sink flow, in which fluid elements not only revolve around the origin, but get sucked into a sink at the origin (actually a line sink along the z-axis). This is illustrated in Fig. 9–15b.
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411 CHAPTER 9
Discussion Other more complicated flows can be obtained by setting f (u, t) to some other function. For any function f (u, t), the flow satisfies the twodimensional, incompressible continuity equation at a given instant in time.
EXAMPLE 9–6
Comparison of Continuity and Volumetric Strain Rate
Recall the volumetric strain rate, defined in Chap. 4. In Cartesian coordinates,
#u #v #w 1 DV " e xx $ e yy $ e zz " $ $ V Dt #x #y #z
SOLUTION We are to show that volumetric strain rate is zero in an incompressible flow, and discuss its physical significance in incompressible and compressible flow. Analysis If the flow is incompressible, Eq. 9–16 applies. More specifically, Eq. 9–17, in Cartesian coordinates, applies. Comparing Eq. 9–17 to Eq. 1,
Time = t1
Volume = V1 (a) Time = t1
for incompressible flow
Thus, volumetric strain rate is zero in an incompressible flow field. In fact, you can define incompressibility by DV/Dt " 0. Physically, as we follow a fluid element, parts of it may stretch while other parts shrink, and the element may translate, distort, and rotate, but its volume remains constant along its entire path through the flow field (Fig. 9–16a). This is true whether the flow is steady or unsteady, as long as it is incompressible. If the flow were compressible, the volumetric strain rate would not be zero, implying that fluid elements may expand in volume (dilate) or shrink in volume as they move around in the flow field (Fig. 9–16b). Specifically, consider Eq. 9–10, an alternative form of the continuity equation for compressible flow. By definition, r " m/V, where m is the mass of a fluid element. For a material element (following the fluid element as it moves through the flow field), m must be constant. Applying some algebra to Eq. 9–10 yields → → 1 Dr V D(m/V) V m DV 1 DV " "& "& " &§% V → r Dt m Dt m V 2 Dt V Dt
1 DV → → ! $%V V Dt
Discussion The result is general—not limited to Cartesian coordinates. It applies to unsteady as well as steady flows.
EXAMPLE 9–7
Time = t2
(1)
Show that volumetric strain rate is zero for incompressible flow. Discuss the physical interpretation of volumetric strain rate for incompressible and compressible flows.
1 DV "0 V Dt
Volume = V2 = V1
Conditions for Incompressible Flow →
→
Consider a steady velocity field given by V " (u, v, w) " a(x 2y $ y 2)i → → $ bxy 2j $ cxk , where a, b, and c are constants. Under what conditions is this flow field incompressible?
Volume = V1 Time = t2 Volume = V2 (b)
FIGURE 9–16 (a) In an incompressible flow field, fluid elements may translate, distort, and rotate, but they do not grow or shrink in volume; (b) in a compressible flow field, fluid elements may grow or shrink in volume as they translate, distort, and rotate.
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412 FLUID MECHANICS
SOLUTION We are to determine a relationship between constants a, b, and c that ensures incompressibility. Assumptions 1 The flow is steady. 2 The flow is incompressible (under certain constraints to be determined). Analysis We apply Eq. 9–17 to the given velocity field, #u #v #w $ $ "0 #x #y #z
→
2axy $ 2bxy " 0
F
F
F 2axy
2bxy
0
Thus to guarantee incompressibility, constants a and b must be equal in magnitude but opposite in sign.
a ! "b
Condition for incompressibility:
Discussion If a were not equal to &b, this might still be a valid flow field, but density would have to vary with location in the flow field. In other words, the flow would be compressible, and Eq. 9–14 would need to be satisfied in place of Eq. 9–17.
9–3
■
THE STREAM FUNCTION
The Stream Function in Cartesian Coordinates
Consider the simple case of incompressible, two-dimensional flow in the xyplane. The continuity equation (Eq. 9–17) in Cartesian coordinates reduces to #u #v $ "0 #x #y
(9–19)
A clever variable transformation enables us to rewrite Eq. 9–19 in terms of one dependent variable (c) instead of two dependent variables (u and v). We define the stream function c as Incompressible, two-dimensional stream function in Cartesian coordinates: u"
#c #y
and
v"&
#c #x
(9–20)
The stream function and the corresponding velocity potential function (Chap. 10) were first introduced by the Italian mathematician Joseph Louis Lagrange (1736–1813). Substitution of Eq. 9–20 into Eq. 9–19 yields #c #2c #2c # #c # ¢ ≤$ ¢& ≤ " & "0 #x #y #y #x #x #y #y #x
which is identically satisfied for any smooth function c(x, y), because the order of differentiation (y then x versus x then y) is irrelevant. You may ask why we chose to put the negative sign on v rather than on u. (We could have defined the stream function with the signs reversed, and continuity would still have been identically satisfied.) The answer is that although the sign is arbitrary, the definition of Eq. 9–20 leads to flow from left to right as c increases in the y-direction, which is usually preferred. Most fluid mechanics books define c in this way, although sometimes c is
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413 CHAPTER 9
defined with the opposite signs (e.g., in the indoor air quality field, Heinsohn and Cimbala, 2003). What have we gained by this transformation? First, as already mentioned, a single variable (c) replaces two variables (u and v)—once c is known, we can generate both u and v via Eq. 9–20, and we are guaranteed that the solution satisfies continuity, Eq. 9–19. Second, it turns out that the stream function has useful physical significance (Fig. 9–17). Namely,
c = c4 c = c3
c = c2
Curves of constant c are streamlines of the flow.
This is easily proven by considering a streamline in the xy-plane, as sketched in Fig. 9–18. Recall from Chap. 4 that along such a streamline, Along a streamline:
→
Streamlines x
F
#c/#x
FIGURE 9–17 Curves of constant stream function represent streamlines of the flow.
#c/#y
where we have applied Eq. 9–20, the definition of c. Thus, #c #c dx $ dy " 0 #x #y
Along a streamline:
y
&v dx $ u dy " 0 F
dy v " dx u
c = c1
(9–21)
But for any smooth function c of two variables x and y, we know by the chain rule of mathematics that the total change of c from point (x, y) to another point (x $ dx, y $ dy) some infinitesimal distance away is dc "
Total change of c:
#c #c dx $ dy #x #y
Streamline
Calculation of the Velocity Field from the Stream Function
A steady, two-dimensional, incompressible flow field in the xy-plane has a stream function given by c " ax 3 $ by $ cx, where a, b, and c are constants: a " 0.50 (m · s)&1, b "&2.0 m/s, and c "&1.5 m/s. (a) Obtain expressions for velocity components u and v. (b) Verify that the flow field satisfies the incompressible continuity equation. (c) Plot several streamlines of the flow in the upper-right quadrant.
SOLUTION For a given stream function, we are to calculate the velocity components, verify incompressibility, and plot flow streamlines. Assumptions 1 The flow is steady. 2 The flow is incompressible (this assumption is to be verified). 3 The flow is two-dimensional in the xy-plane, implying that w " 0 and neither u nor v depend on z. Analysis (a) We use Eq. 9–20 to obtain expressions for u and v by differentiating the stream function, u"
#c "b #y
and
v"&
#c " "3ax2 " c #x
(b) Since u is not a function of x, and v is not a function of y, we see immediately that the two-dimensional, incompressible continuity equation (Eq. 9–19) is satisfied. In fact, since c is smooth in x and y, the two-dimensional,
→
V →
dr
(9–22)
By comparing Eq. 9–21 to Eq. 9–22 we see that dc " 0 along a streamline; thus we have proven the statement that c is constant along streamlines.
EXAMPLE 9–8
Point (x + dx, y + dy)
v
dy dx y
u
Point (x, y) x
FIGURE 9–18 → Arc length dr " (dx, dy) and local → velocity vector V " (u, v) along a two-dimensional streamline in the xy-plane.
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414 FLUID MECHANICS 10 m/s
Scale for velocity vectors: 5 –10 4
60 –7.5
3 y, m
50 40
2
c = –5
m2/s
30
1
20 10
c=0
0
5 –1 0
1
2
3 x, m
4
FIGURE 9–19 Streamlines for the velocity field of Example 9–8; the value of constant c is indicated for each streamline, and velocity vectors are shown at four locations.
5
incompressible continuity equation in the xy-plane is automatically satisfied by the very definition of c. We conclude that the flow is indeed incompressible. (c) To plot streamlines, we solve the given equation for either y as a function of x and c, or x as a function of y and c. In this case, the former is easier, and we have
Equation for a streamline:
y"
c $ ax 3 $ cx b
This equation is plotted in Fig. 9–19 for several values of c, and for the provided values of a, b, and c. The flow is nearly straight down at large values of x, but veers upward for x ! 1 m. Discussion You can verify that v " 0 at x " 1 m. In fact, v is negative for x # 1 m and positive for x ! 1 m. The direction of the flow can also be determined by picking an arbitrary point in the flow, say (x " 3 m, y " 4 m), and calculating the velocity there. We get u " $2.0 m/s and v " $12.0 m/s at this point, either of which shows that fluid flows to the lower left in this region of the flow field. For clarity, the velocity vector at this point is also plotted in Fig. 9–19; it is clearly parallel to the streamline near that point. Velocity vectors at three other points are also plotted.
EXAMPLE 9–9
Calculation of Stream Function for a Known Velocity Field
Consider a steady, two-dimensional, incompressible velocity field with u " ax % b and v " $ay % cx, where a, b, and c are constants: a " 0.50 s$1, b " 1.5 m/s, and c " 0.35 s$1. Generate an expression for the stream function and plot some streamlines of the flow in the upper-right quadrant.
SOLUTION For a given velocity field we are to generate an expression for c and plot several streamlines for given values of constants a, b, and c. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the xy-plane, implying that w " 0 and neither u nor v depend on z. Analysis We start by picking one of the two parts of Eq. 9–20 that define the stream function (it doesn’t matter which part we choose—the solution will be identical). 'c " u " ax % b 'y Next we integrate with respect to y, noting that this is a partial integration, so we add an arbitrary function of the other variable, x, rather than a constant of integration,
c " axy % by % g(x)
(1)
Now we choose the other part of Eq. 9–20, differentiate Eq. 1, and rearrange as follows:
v"$
'c " $ay $ g&(x) 'x
(2)
where g&(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, the equation given in the
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415 CHAPTER 9
problem statement and Eq. 2. We equate these and integrate with respect to x to find g (x),
x2 v " &ay $ cx " &ay & g)(x) → g)(x) " &cx → g(x) " &c $ C (3) 2 Note that here we have added an arbitrary constant of integration C since g is a function of x only. Finally, substituting Eq. 3 into Eq. 1 yields the final expression for c,
Solution:
C ! axy # by " c
x2 #C 2
c $ cx 2( 2 ax $ b
Equation for streamlines:
y"
(5)
For the given values of constants a, b, and c, we plot Eq. 5 for several values of c in Fig. 9–20; these curves of constant c are streamlines of the flow. From Fig. 9–20 we see that this is a smoothly converging flow in the upperright quadrant. Discussion It is always good to check your algebra. In this example, you should substitute Eq. 4 into Eq. 9–20 to verify that the correct velocity components are obtained.
14 12
4
10 8
3 6
y, m
(4)
To plot the streamlines, we note that Eq. 4 represents a family of curves, one unique curve for each value of the constant (c & C). Since C is arbitrary, it is common to set it equal to zero, although it can be set equal to any desired value. For simplicity we set C " 0 and solve Eq. 4 for y as a function of x, yielding
16
5
2
c = 4 m2/s
1
2 0 –2
0
–4
–6
–1 0
1
2
3 x, m
4
5
FIGURE 9–20 Streamlines for the velocity field of Example 9–9; the value of constant c is indicated for each streamline.
There is another physically significant fact about the stream function: The difference in the value of c from one streamline to another is equal to the volume flow rate per unit width between the two streamlines.
This statement is illustrated in Fig. 9–21. Consider two streamlines, c1 and c2, and imagine two-dimensional flow in the xy-plane, of unit width into the page (1 m in the &z-direction). By definition, no flow can cross a streamline. Thus, the fluid that happens to occupy the space between these two streamlines remains confined between the same two streamlines. It follows that the mass flow rate through any cross-sectional slice between the streamlines is the same at any instant in time. The cross-sectional slice can be any shape, provided that it starts at streamline 1 and ends at streamline 2. In Fig. 9–21, for example, slice A is a smooth arc from one streamline to the other while slice B is wavy. For steady, incompressible, two-dimensional . flow in the xy-plane, the volume flow rate V between the two streamlines (per unit width) must therefore be a constant. If the two streamlines spread apart, as they do from cross-sectional slice A to cross-sectional slice B, the average velocity between the two streamlines decreases . accordingly, such . that the volume flow rate remains the same (VA " VB). In Fig. 9–19 of Example 9–8, velocity vectors at four locations in the flow field between streamlines c " 0 m2/s and c " 5 m2/s are plotted. You can clearly see that as the streamlines diverge from each other, the velocity vector decays in
c = c2
Streamline 2
⋅ ⋅ VB = VA ⋅ VA
B
A
y c = c1 Streamline 1
x
FIGURE 9–21 For two-dimensional streamlines in the xy-plane, the volume flow rate . V per unit width between two streamlines is the same through any cross-sectional slice.
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416 FLUID MECHANICS Streamline 2
magnitude. Likewise, when streamlines converge, the average velocity between them must increase. We prove the given statement mathematically by considering a control volume bounded by the two streamlines of Fig. 9–21 and by cross-sectional slice A and cross-sectional slice B (Fig. 9–22). An infinitesimal length ds along → slice B is illustrated in Fig. 9–22a, along with its unit normal vector n . A magnified view of this region is sketched in Fig. 9–22b for clarity. As shown, the two components of ds are dx and dy; thus the unit normal vector is
c = c2
→
V
ds
→
n
CV B
A
y c = c1
→
x
Streamline 1 (a)
CV
n"
dy → dx → i & j ds ds
The volume flow rate per unit width through segment ds of the control surface is
Control surface
# → → → → dy → dx → dV " V % n dA " (ui $ v j ) % ¢ i & j ≤ ds ds ds
→
V
(9–23)
F ds
v dy/ds
ds
where dA " ds times 1 " ds, where the 1 indicates a unit width into the page, regardless of the unit system. When we expand the dot product of Eq. 9–23 and apply Eq. 9–20, we get
u dx ds
dy
# #c #c dV " u dy & v dx " dy $ dx " dc #y #x
→
n
dx
y x (b)
FIGURE 9–22 (a) Control volume bounded by streamlines c1 and c2 and slices A and B in the xy-plane; (b) magnified view of the region around infinitesimal length ds.
y x c=7 c=6
(9–24)
We find the total volume flow rate through cross-sectional slice B by integrating Eq. 9–24 from streamline 1 to streamline 2, # VB "
#
! V % n dA " ! dV " ! →
B
→
B
c"c2
dc " c2 & c1
(9–25)
c"c1
Thus, the volume flow rate per unit width through slice B is equal to the difference between the values of the two stream functions that bound slice B. Now consider the entire control volume of Fig. 9–22a. Since we know that no flow crosses the streamlines, conservation of mass demands that the volume flow rate into the control volume through slice A be identical to the volume flow rate out of the control volume through slice B. Finally, since we can choose a cross-sectional slice of any shape or location between the two streamlines, the statement is proven. When dealing with stream functions, the direction of flow is obtained by what we might call the “left-side convention.” Namely, if you are looking down the z-axis at the xy-plane (Fig. 9–23) and are moving in the direction of the flow, the stream function increases to your left. The value of c increases to the left of the direction of flow in the xy-plane.
c=5
FIGURE 9–23 Illustration of the “left-side convention.” In the xy-plane, the value of the stream function always increases to the left of the flow direction.
In Fig. 9–23, for example, the stream function increases to the left of the flow direction, regardless of how much the flow twists and turns. Notice also that when the streamlines are far apart (lower right of Fig. 9–23), the magnitude of velocity (the fluid speed) in that vicinity is small relative to the speed in locations where the streamlines are close together (middle region of Fig. 9–23). This is easily explained by conservation of mass. As the streamlines converge, the cross-sectional area between them decreases, and the velocity must increase to maintain the flow rate between the streamlines.
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417 CHAPTER 9
FIGURE 9–24 Streaklines produced by Hele–Shaw flow over an inclined plate. The streaklines model streamlines of potential flow (Chap. 10) over a two-dimensional inclined plate of the same cross-sectional shape. Courtesy Howell Peregrine, School of Mathematics, University of Bristol. Used by permission.
EXAMPLE 9–10
Relative Velocity Deduced from Streamlines
Hele–Shaw flow is produced by forcing a liquid through a thin gap between parallel plates. An example of Hele–Shaw flow is provided in Fig. 9–24 for flow over an inclined plate. Streaklines are generated by introducing dye at evenly spaced points upstream of the field of view. Since the flow is steady, the streaklines are coincident with streamlines. The fluid is water and the glass plates are 1.0 mm apart. Discuss how you can tell from the streamline pattern whether the flow speed in a particular region of the flow field is (relatively) large or small.
SOLUTION For the given set of streamlines, we are to discuss how we can tell the relative speed of the fluid. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow models two-dimensional potential flow in the xy-plane. Analysis When equally spaced streamlines of a stream function spread away from each other, it indicates that the flow speed has decreased in that region. Likewise, if the streamlines come closer together, the flow speed has increased in that region. In Fig. 9–24 we infer that the flow far upstream of the plate is straight and uniform, since the streamlines are equally spaced. The fluid decelerates as it approaches the underside of the plate, especially near the stagnation point, as indicated by the wide gap between streamlines. The flow accelerates rapidly to very high speeds around the sharp corners of the plate, as indicated by the tightly spaced streamlines. Discussion The streaklines of Hele–Shaw flow turn out to be similar to those of potential flow, which is discussed in Chap. 10.
EXAMPLE 9–11
Volume Flow Rate Deduced from Streamlines
Water is sucked through a narrow slot on the bottom wall of a water channel. The water in the channel flows from left to right at uniform velocity V " 1.0 m/s. The slot is perpendicular to the xy-plane, and runs along the zaxis across the entire channel, which is w " 2.0 m wide. The flow is thus approximately two-dimensional in the xy-plane. Several streamlines of the flow are plotted and labeled in Fig. 9–25.
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418 FLUID MECHANICS 2
FIGURE 9–25 Streamlines for free-stream flow along a wall with a narrow suction slot; streamline values are shown in units of m2/s; the thick streamline is the dividing streamline. The direction of the velocity vector at point A is determined by the left-side convention.
2.0 1.8
1.5
1.6
A 0.4
y, m 1
1.0
0.5
0.4
1.2
0.8 0.6 0.2
0 –3
–2
–1
⋅ V w
1
2
x, m
The thick streamline in Fig. 9–25 is called the dividing streamline because it divides the flow into two parts. Namely, all the water below this dividing streamline gets sucked into the slot, while all the water above the dividing streamline continues on its way downstream. What is the volume flow rate of water being sucked through the slot? Estimate the magnitude of the velocity at point A.
SOLUTION For the given set of streamlines, we are to determine the volume flow rate through the slot and estimate the fluid speed at a point. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the xy-plane. 4 Friction along the bottom wall is neglected. Analysis By Eq. 9–25, the volume flow rate per unit width between the bottom wall (cwall " 0) and the dividing streamline (cdividing " 1.0 m2/s) is # V " cdividing & cwall " (1.0 & 0) m2/s " 1.0 m2/s w
All of this flow must go through the slot. Since the channel is 2.0 m wide, the total volume flow rate through the slot is
# # V V " w " (1.0 m2/s)(2.0 m) " 2.0 m3/s w
To estimate the speed at point A, we measure the distance d between the two streamlines that enclose point A. We find that streamline 1.8 is about 0.21 m away from streamline 1.6 in the vicinity of point A. The volume flow rate per unit width (into the page) between these two streamlines is equal to the difference in value of the stream function. We can thus estimate the speed at point A,
# # V 1V 1 1 " (c1.8 & c1.6) " VA # " (1.8 & 1.6) m2/s " 0.95 m/s wd d w d 0.21 m
Our estimate agrees very well with the known free-stream speed (1.0 m/s), indicating that the fluid in the vicinity of point A flows at nearly the same speed as the free-stream flow, but points slightly downward. Discussion The streamlines of Fig. 9–25 were generated by superposition of a uniform stream and a line sink, assuming irrotational (potential) flow. We discuss such superposition in Chap. 10.
3
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419 CHAPTER 9
The Stream Function in Cylindrical Coordinates
For two-dimensional flow, we can also define the stream function in cylindrical coordinates, which is more convenient for many problems. Note that by two-dimensional we mean that there are only two relevant independent spatial coordinates—with no dependence on the third component. There are two possibilities. The first is planar flow, just like that of Eqs. 9–19 and 9–20, but in terms of (r, u) and (ur, uu) instead of (x, y) and (u, v) (see Fig. 9–10a). In this case, there is no dependence on coordinate z. We simplify the incompressible continuity equation, Eq. 9–18, for two-dimensional planar flow in the ru-plane, #(ru r) #(u u) $ !0 #r #u
(9–26)
We define the stream function as follows: Incompressible, planar stream function in cylindrical coordinates: ur !
1 #c r #u
and
uu ! "
#c #r
(9–27)
We note again that the signs are reversed in some textbooks. You can substitute Eq. 9–27 into Eq. 9–26 to convince yourself that Eq. 9–26 is identically satisfied for any smooth function c(r, u), since the order of differentiation (r then u versus u then r) is irrelevant for a smooth function. The second type of two-dimensional flow in cylindrical coordinates is axisymmetric flow, in which r and z are the relevant spatial variables, ur and uz are the nonzero velocity components, and there is no dependence on u (Fig. 9–26). Examples of axisymmetric flow include flow around spheres, bullets, and the fronts of many objects like torpedoes and missiles, which would be axisymmetric everywhere if not for their fins. For incompressible axisymmetric flow, the continuity equation is 1 #(ru r) #(u z) $ !0 r #r #z
(9–28)
The stream function c is defined such that it satisfies Eq. 9–28 exactly, provided of course that c is a smooth function of r and z, Incompressible, axisymmetric stream function in cylindrical coordinates: ur ! "
1 #c r #z
and
uz !
1 #c r #r
(9–29)
We also note that there is another way to describe axisymmetric flows, namely, by using Cartesian coordinates (x, y) and (u, v), but forcing coordinate x to be the axis of symmetry. This can lead to confusion because the equations of motion must be modified accordingly to account for the axisymmetry. Nevertheless, this is often the approach used in CFD codes. The advantage is that after one sets up a grid in the xy-plane, the same grid can be used for both planar flow (flow in the xy-plane with no z-dependence) and axisymmetric flow (flow in the xy-plane with rotational symmetry about the x-axis). We do not discuss the equations for this alternative description of axisymmetric flows.
y z
r u
x
ur r
uz
Rotational symmetry Axisymmetric body z
FIGURE 9–26 Flow over an axisymmetric body in cylindrical coordinates with rotational symmetry about the z-axis; neither the geometry nor the velocity field depend on u, and uu ! 0.
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420 FLUID MECHANICS
EXAMPLE 9–12
Stream Function in Cylindrical Coordinates
Consider a line vortex, defined as steady, planar, incompressible flow in which the velocity components are ur " 0 and uu " K/r, where K is a constant. This flow is represented in Fig. 9–15a. Derive an expression for the stream function c(r, u), and prove that the streamlines are circles.
SOLUTION For a given velocity field in cylindrical coordinates, we are to derive an expression for the stream function. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is planar in the ru-plane. Analysis We use the definition of stream function given by Eq. 9–27. We can choose either component to start with; we choose the tangential component, #c K " &u u " & r #r
→
c " &K ln r $ f (u)
(1)
Now we use the other component of Eq. 9–27,
ur "
2
f )(u) " 0
0.5
y
4
(3)
Finally, we see from Eq. 3 that curves of constant c are produced by setting r to a constant value. Since curves of constant r are circles by definition, streamlines (curves of constant c) must therefore be circles about the origin, as in Fig. 9–15a. For given values of C and c, we solve Eq. 3 for r to plot the streamlines,
8 10 –1
f (u) " C
C ! "K ln r # C
Solution:
–0.5
-1
→
where C is an arbitrary constant of integration. Equation 1 is thus
6
0
(2)
where the prime denotes a derivative with respect to u. By equating ur from the given information to Eq. 2, we see that
22
c = 0 m2/s
1
1 #c 1 " f )(u) r #u r
–0.5
12 0 x
0.5
14
Equation for streamlines: 1
FIGURE 9–27 Streamlines for the velocity field of Example 9–12, with K " 10 m2/s and C " 0; the value of constant c is indicated for several streamlines.
r " e & (c& C)/K
(4)
m2/s
For K " 10 and C " 0, streamlines from c " 0 to 22 are plotted in Fig. 9–27. Discussion Notice that for a uniform increment in the value of c, the streamlines get closer and closer together near the origin as the tangential velocity increases. This is a direct result of the statement that the difference in the value of c from one streamline to another is equal to the volume flow rate per unit width between the two streamlines.
The Compressible Stream Function*
We extend the stream function concept to steady, compressible, two-dimensional flow in the xy-plane. The compressible continuity equation (Eq. 9–14) in Cartesian coordinates reduces to the following for steady twodimensional flow: #(ru) #(rv) $ "0 #x #y * This section can be skipped without loss of continuity.
(9–30)
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421 CHAPTER 9
We introduce a compressible stream function, which we denote as cr, Steady, compressible, two-dimensional stream function in Cartesian coordinates: ru "
#cr
and
#y
rv " &
#cr
(9–31)
#x
By definition, cr of Eq. 9–31 satisfies Eq. 9–30 exactly, provided that cr is a smooth function of x and y. Many of the features of the compressible stream function are the same as those of the incompressible c as discussed previously. For example, curves of constant cr are still streamlines. However, the difference in cr from one streamline to another is mass flow rate per unit width rather than volume flow rate per unit width. Although not as popular as its incompressible counterpart, the compressible stream function finds use in some commercial CFD codes.
9–4
■
CONSERVATION OF LINEAR MOMENTUM— CAUCHY’S EQUATION
Through application of the Reynolds transport theorem (Chap. 4), we have the general expression for conservation of linear momentum as applied to a control volume, aF" →
!
→
rg dV $
CV
!
→
sij % n dA "
CS
!
CV
# → (rV) dV $ #t
!
→ →
→
(rV )V % n dA
(9–32)
CS
where sij is the stress tensor introduced in Chap. 6. Components of sij on the positive faces of an infinitesimal rectangular control volume are shown in Fig. 9–28. Equation→9–32 applies to both fixed and moving control volumes, provided that V is the absolute velocity (as seen from a fixed observer). For the special case of flow with well defined inlets and outlets, Eq. 9–32 can be simplified as follows: a F " a Fbody $ a Fsurface " →
→
→
# → #→ #→ (rV ) dV $ a bmV & a bmV #t out in CV
!
syy
(9–33)
syx
→
where V in the last two terms is taken as the average velocity at an inlet or outlet, and b is the momentum flux correction factor (Chap. 6). In words, the total force acting on the control volume is equal to the rate at which momentum changes within the control volume plus the rate at which momentum flows out of the control volume minus the rate at which momentum flows into the control volume. Equation 9–33 applies to any control volume, regardless of its size. To generate a differential equation for conservation of linear momentum, we imagine the control volume shrinking to infinitesimal size. In the limit, the entire control volume shrinks to a point in the flow (Fig. 9–2). We take the same approach here as we did for conservation of mass; namely, we show more than one way to derive the differential form of conservation of linear momentum.
Derivation Using the Divergence Theorem
The most straightforward (and most elegant) way to derive the differential form of conservation of momentum is to apply the divergence theorem of Eq. 9–3. A more general form of the divergence theorem applies not only to vectors, but to other quantities as well, such as tensors, as illustrated in Fig.
syz sxy sxx
szy dy
szx
szz
sxz dz
dx
FIGURE 9–28 Positive components of the stress tensor in Cartesian coordinates on the positive (right, top, and front) faces of an infinitesimal rectangular control volume. The blue dots indicate the center of each face. Positive components on the negative (left, bottom, and back) faces are in the opposite direction of those shown here.
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422 FLUID MECHANICS
The Extended Divergence Theorem →
∆
!V
•
9–29. Specifically, if we replace G in the extended divergence theorem of → → ij Fig. 9–29 with the quantity (rV )V , a second-order tensor, the last term in Eq. 9–32 becomes
Gij dV = "A Gij • n dA →
!
→ →
→
(rV )V % n dA "
CS
!
→
→→
§ % (rV V ) dV
(9–34)
CV
→→
FIGURE 9–29 An extended form of the divergence theorem is useful not only for vectors, but also for tensors. In the equation, Gij is a second-order tensor, V is a volume, and A is the surface area that encloses and defines the volume.
where V V is a vector product called the outer product of the velocity vector with itself. (The outer product of two vectors is not the same as the inner or dot product, nor is it the same as the cross product of the two vectors.) Similarly, if we replace Gij in Fig. 9–29 by the stress tensor sij, the second term on the left-hand side of Eq. 9–32 becomes
!
CS
CV
ay
∂ (rV ) + ∂t
→
∆
→
→→
•
(rV V ) =
→
rg +
Cauchy’s equation: →
∆
y’s Cauch
→
§ % sij dV
(9–35)
CV
→ →→ → # → → c (rV ) $ § % (rV V ) & rg & § % sijd dV " 0 #t
(9–36)
Finally, we argue that Eq. 9–36 must hold for any control volume regardless of its size or shape. This is possible only if the integrand (enclosed by square brackets) is identically zero. Hence, we have a general differential equation for conservation of linear momentum, known as Cauchy’s equation,
of the D
on Equati
!
Thus, the two surface integrals of Eq. 9–32 become volume integrals by applying Eqs. 9–34 and 9–35. We combine and rearrange the terms, and rewrite Eq. 9–32 as
!
n Equatio
→
sij % n dA "
•
s ij
FIGURE 9–30 Cauchy’s equation is a differential form of the law of conservation of linear momentum. It applies to any type of fluid.
→ →→ → # → → (rV ) $ § % (rV V ) " rg $ § % sij #t
(9–37)
Equation 9–37 is named in honor of the French engineer and mathematician Augustin Louis de Cauchy (1789–1857). It is valid for compressible as well as incompressible flow since we have not made any assumptions about incompressibility. It is valid at any point in the flow domain (Fig. 9–30). Note that Eq. 9–37 is a vector equation, and thus represents three scalar equations, one for each coordinate axis in three-dimensional problems.
Derivation Using an Infinitesimal Control Volume
We derive Cauchy’s equation a second way, using an infinitesimal control volume on which we apply conservation of linear momentum (Eq. 9–33). We consider the same box-shaped control volume we used to derive the continuity equation (Fig. 9–3). At the center of the box, as previously, we define the density as r and the velocity components as u, v, and w. We also we define the stress tensor as sij at the center of the box. For simplicity, → equal to consider the x-component of Eq. 9–33, obtained by setting F a → its x-component, a Fx, and V equal to its x-component, u. This not only simplifies the diagrams, but enables us to work with a scalar equation, namely, a Fx " a Fx, body $ a Fx, surface "
# # # (ru) dV $ a bmu & a bmu (9–38) #t out in CV
!
As the control volume shrinks to a point, the first term on the right-hand side of Eq. 9–38 becomes
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423 CHAPTER 9 y arvu +
x z
∂(rvu) dy b dx dz ∂y 2 arwu –
∂(rwu) dz b dx dy ∂z 2
dy ∂(ruu) dx aruu – b dy dz ∂x 2
aruu +
∂(ruu) dx b dy dz ∂x 2
dz
∂(rwu) dz arwu + b dx dy ∂z 2
dx arvu –
∂(rvu) dy b dx dz ∂y 2
FIGURE 9–31 Inflow and outflow of the x-component of linear momentum through each face of an infinitesimal control volume; the blue dots indicate the center of each face.
Rate of change of x-momentum within the control volume:
"
$ $ (ru) dV ! (ru) dx dy dz $t $t CV
(9–39)
since the volume of the differential element is dx dy dz. We apply first-order truncated Taylor series expansions at locations away from the center of the control volume to approximate the inflow and outflow of momentum in the x-direction. Figure 9–31 shows these momentum fluxes at the center point of each of the six faces of the infinitesimal control volume. Only the normal velocity component at each face needs to be considered, since the tangential velocity components contribute no mass flow out of (or into) the face, and hence no momentum flow through the face either. By summing all the outflows and subtracting all the inflows shown in Fig. 9–31, we obtain an approximation for the last two terms of Eq. 9–38, →
(9–40)
where b is set equal to one at all faces, consistent with our first-order approximation. Next, we sum all the forces acting on our infinitesimal control volume in the x-direction. As was done in Chap. 6, we need to consider both body forces and surface forces. Gravity force (weight) is the only body force we take into account. For the general case in which the coordinate system may not be aligned with the z-axis (or with any coordinate axis for that matter), as sketched in Fig. 9–32, the gravity vector is written as →
→
→
→
g ! gx i " gy j " gz k
Thus, in the x-direction, the body force on the control volume is a Fx, body ! a Fx, gravity ! rgx dx dy dz
dy g
Net outflow of x-momentum through the control surface:
$ $ $ # # a bmu # a bmu ! ¢$x (ruu) " $y (rvu) " $z (rwu)≤ dx dy dz out in
dx
(9–41)
Next we consider the net surface force in the x-direction. Recall that stress tensor sij has dimensions of force per unit area. Thus, to obtain a
dz y x z
→
Fgravity
FIGURE 9–32 The gravity vector is not necessarily aligned with any particular axis, in general, and there are three components of the body force acting on an infinitesimal fluid element.
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424 FLUID MECHANICS aszx –
FIGURE 9–33 Sketch illustrating the surface forces acting in the x-direction due to the appropriate stress tensor component on each face of the differential control volume; the blue dots indicate the center of each face.
asxx –
∂szx dz b dx dy ∂z 2
∂sxx dx b dy dz ∂x 2
dx asyx +
∂syx dy b dx dz ∂y 2
asxx +
dy
∂s dy asyx – yx b dx dz ∂y 2
∂sxx dx b dy dz ∂x 2
dz
y x
aszx +
∂szx dz b dx dy ∂z 2
z
force, we must multiply each stress component by the surface area of the face on which it acts. We need to consider only those components that point in the x- (or &x-) direction. (The other components of the stress tensor, although they may be nonzero, do not contribute to a net force in the xdirection.) Using truncated Taylor series expansions, we sketch all the surface forces that contribute to a net x-component of surface force acting on our differential fluid element (Fig. 9–33). Summing all the surface forces illustrated in Fig. 9–33, we obtain an approximation for the net surface force acting on the differential fluid element in the x-direction, # # # a Fx, surface # ¢#x sxx $ #y syx $ #z szx≤ dx dy dz
(9–42)
We now substitute Eqs. 9–39 through 9–42 into Eq. 9–38, noting that the volume of the differential element of fluid, dx dy dz, appears in all terms and can be eliminated. After some rearrangement we obtain the differential form of the x-momentum equation, #(ru) #(ruu) #(rvu) #(rwu) # # # s $ s $ s $ $ $ " rgx $ #t #x #y #z #x xx #y yx #z zx
(9–43)
In similar fashion, we generate differential forms of the y- and z-momentum equations, #(rv) #(ruv) #(rvv) #(rwv) # # # $ $ $ " rgy $ s $ s $ s #t #x #y #z #x xy #y yy #z zy
(9–44)
#(rw) #(ruw) #(rvw) #(rww) # # # $ $ $ " rgz $ s $ s $ s #t #x #y #z #x xz #y yz #z zz
(9–45)
and
respectively. Finally, we combine Eqs. 9–43 through 9–45 into one vector equation, Cauchy’s equation:
→ →→ → # → → (rV ) $ § % (rV V ) " rg $ § % sij #t
This equation is identical to Cauchy’s equation (Eq. 9–37); thus we confirm that our derivation using the differential fluid element yields the same result
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425 CHAPTER 9 →→
as our derivation using the divergence theorem. Note that the product V V is a second-order tensor (Fig. 9–34). Outer Product:
Alternative Form of Cauchy’s Equation
uu uv uw VV = vu vv vw wu wv ww
Applying the product rule to the first term on the left side of Eq. 9–37, we get
→→
→
→ #r # → #V (rV ) " r $V #t #t #t
(9–46)
The second term of Eq. 9–37 can be written as →
→→
→→
→
→
→ →
§ % (rV V ) " V § % (rV ) $ r(V % §)V
(9–47) →→
Thus we have eliminated the second-order tensor represented by V V . After some rearrangement, substitution of Eqs. 9–46 and 9–47 into Eq. 9–37 yields →
FIGURE 9–34 → The outer product of vector V " (u, v, w) with itself is a secondorder tensor. The product shown is in Cartesian coordinates and is illustrated as a nine-component matrix.
→ #r → → → → → → #V → $ V c $ § % (rV)d $ r( V % §)V " rg $ § % sij r #t #t
But the expression in square brackets in this equation is identically zero by the continuity equation, Eq. 9–5. By combining the remaining two terms on the left side, we write Alternative form of Cauchy’s equation: →
→
→ → → → DV #V → " rg $ § % sij rc $ (V % §)Vd " r #t Dt
(9–48)
where we have recognized the expression in square brackets as the material acceleration—the acceleration following a fluid particle (see Chap. 4).
Derivation Using Newton’s Second Law
We derive Cauchy’s equation by yet a third method. Namely, we take the differential fluid element as a material element instead of a control volume. In other words, we think of the fluid within the differential element as a tiny system of fixed identity, moving→with the flow (Fig. 9–35). The acceleration → of this fluid element is a " DV /Dt by definition of the material acceleration. By Newton’s second law applied to a material element of fluid, → DV DV → a F " ma " m Dt " r dx dy dz Dt →
→
ΣF
(9–49)
→
a
→
→ DV → " rg $ § % sij Dt
x z
→
At the instant in time represented in Fig. 9–35, the net force on the differential fluid element is found in the same way as that calculated earlier on the differential control volume. Thus the total force acting on the fluid element is the sum of Eqs. 9–41 and 9–42, extended to vector form. Substituting these into Eq. 9–49 and dividing by dx dy dz, we once again generate the alternative form of Cauchy’s equation, r
y
(9–50)
Equation 9–50 is identical to Eq. 9–48. In hindsight, we could have started with Newton’s second law from the beginning, avoiding some algebra. Nevertheless, derivation of Cauchy’s equation by three methods certainly boosts our confidence in the validity of the equation!
dy dz dx Streamline
FIGURE 9–35 If the differential fluid element is a material element, it moves with the flow and Newton’s second law applies directly.
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426 FLUID MECHANICS
We must be very careful when expanding the last term of Eq. 9–50, which is the divergence of a second-order tensor. In Cartesian coordinates, the three components of Cauchy’s equation are x-component:
r
$sxx $syx $szx Du ! rgx # # # Dt $x $y $z
(9–51a)
y-component:
r
$sxy $syy $szy Dv ! rgy # # # Dt $x $y $z
(9–51b)
z-component:
r
$sxz $syz $szz Dw ! rgz # # # Dt $x $y $z
(9–51c)
We conclude this section by noting that we cannot solve any fluid mechanics problems using Cauchy’s equation by itself (even when combined with continuity). The problem is that the stress tensor sij needs to be expressed in terms of the primary unknowns in the problem, namely, density, pressure, and velocity. This is done for the most common type of fluid in Section 9–5.
9–5
■
THE NAVIER–STOKES EQUATION
Introduction
y x
P
z
P P
P dy
Cauchy’s equation (Eq. 9–37 or its alternative form Eq. 9–48) is not very useful to us as is, because the stress tensor sij contains nine components, six of which are independent (because of symmetry). Thus, in addition to density and the three velocity components, there are six additional unknowns, for a total of 10 unknowns. (In Cartesian coordinates the unknowns are r, u, v, w, sxx, sxy, sxz, syy, syz, and szz). Meanwhile, we have discussed only four equations so far—continuity (one equation) and Cauchy’s equation (three equations). Of course, to be mathematically solvable, the number of equations must equal the number of unknowns, and thus we need six more equations. These equations are called constitutive equations, and they enable us to write the components of the stress tensor in terms of the velocity field and pressure field. The first thing we do is separate the pressure stresses and the viscous stresses. When a fluid is at rest, the only stress acting at any surface of any fluid element is the local hydrostatic pressure P, which always acts inward and normal to the surface (Fig. 9–36). Thus, regardless of the orientation of the coordinate axes, for a fluid at rest the stress tensor reduces to Fluid at rest:
P dz dx
P
FIGURE 9–36 For fluids at rest, the only stress on a fluid element is the hydrostatic pressure, which always acts inward and normal to any surface.
sxx sxy sxz "P 0 0 sij ! £syx syy syz ≥ ! £ 0 "P 0 ≥ szx szy szz 0 0 "P
(9–52)
Hydrostatic pressure P in Eq. 9–52 is the same as the thermodynamic pressure with which we are familiar from our study of thermodynamics. P is related to temperature and density through some type of equation of state (e.g., the ideal gas law). As a side note, this further complicates a compressible fluid flow analysis because we introduce yet another unknown, namely, temperature T. This new unknown requires another equation—the differential form of the energy equation—which is not discussed in this text.
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427 CHAPTER 9
When a fluid is moving, pressure still acts inwardly normal, but viscous stresses may also exist. We generalize Eq. 9–52 for moving fluids as Moving fluids: sxx sij " £syx szx
sxy syy szy
&P sxz syz ≥ " £ 0 szz 0
0 0 txx &P 0≥ $ £tyx 0 &P tzx
txy tyy tzy
txz tyz ≥ tzz
(9–53)
where we have introduced a new tensor, tij, called the viscous stress tensor or the deviatoric stress tensor. Mathematically, we have not helped the situation because we have replaced the six unknown components of sij with six unknown components of tij, and have added another unknown, pressure P. Fortunately, however, there are constitutive equations that express tij in terms of the velocity field and measurable fluid properties such as viscosity. The actual form of the constitutive relations depends on the type of fluid, as discussed shortly. As a side note, there are some subtleties associated with the pressure in Eq. 9–53. If the fluid is incompressible, we have no equation of state (it is replaced by the equation r " constant), and we can no longer define P as the thermodynamic pressure. Instead, we define P in Eq. 9–53 as the mechanical pressure, Mechanical pressure:
1 Pm " & (sxx $ syy $ szz) 3
(9–54)
We see from Eq. 9–54 that mechanical pressure is the mean normal stress acting inwardly on a fluid element. It is therefore also called mean pressure by some authors. Thus, when dealing with incompressible fluid flows, pressure variable P is always interpreted as the mechanical pressure Pm. For compressible flow fields however, pressure P in Eq. 9–53 is the thermodynamic pressure, but the mean normal stress felt on the surfaces of a fluid element is not necessarily the same as P (pressure variable P and mechanical pressure Pm are not necessarily equivalent). You are referred to Panton (1996) or Kundu (1990) for a more detailed discussion of mechanical pressure.
Newtonian versus Non-Newtonian Fluids
The study of the deformation of flowing fluids is called rheology; the rheological behavior of various fluids is sketched in Fig. 9–37. In this text, we concentrate on Newtonian fluids, defined as fluids for which the shear stress is linearly proportional to the shear strain rate. Newtonian fluids (stress proportional to strain rate) are analogous to elastic solids (Hooke’s law: stress proportional to strain). Many common fluids, such as air and other gases, water, kerosene, gasoline, and other oil-based liquids, are Newtonian fluids. Fluids for which the shear stress is not linearly related to the shear strain rate are called non-Newtonian fluids. Examples include slurries and colloidal suspensions, polymer solutions, blood, paste, and cake batter. Some non-Newtonian fluids exhibit a “memory”—the shear stress depends not only on the local strain rate, but also on its history. A fluid that returns (either fully or partially) to its original shape after the applied stress is released is called viscoelastic.
Bingham plastic
Shear stress
Shear thinning Yield stress
Newtonian Shear thickening
Shear strain rate
FIGURE 9–37 Rheological behavior of fluids—shear stress as a function of shear strain rate.
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428 FLUID MECHANICS I think he means quicksand.
Help!
I fell into a dilatant fluid!
?
FIGURE 9–38 When an engineer falls into quicksand (a dilatant fluid), the faster he tries to move, the more viscous the fluid becomes.
Some non-Newtonian fluids are called shear thinning fluids or pseudoplastic fluids, because the more the fluid is sheared, the less viscous it becomes. A good example is paint. Paint is very viscous when poured from the can or when picked up by a paintbrush, since the shear rate is small. However, as we apply the paint to the wall, the thin layer of paint between the paintbrush and the wall is subjected to a large shear rate, and it becomes much less viscous. Plastic fluids are those in which the shear thinning effect is extreme. In some fluids a finite stress called the yield stress is required before the fluid begins to flow at all; such fluids are called Bingham plastic fluids. Certain pastes such as acne cream and toothpaste are examples of Bingham plastic fluids. If you hold the tube upside down, the paste does not flow, even though there is a nonzero stress due to gravity. However, if you squeeze the tube (greatly increasing the stress), the paste flows like a very viscous fluid. Other fluids show the opposite effect and are called shear thickening fluids or dilatant fluids; the more the fluid is sheared, the more viscous it becomes. The best example is quicksand, a thick mixture of sand and water. As we all know from Hollywood movies, it is easy to move slowly through quicksand, since the viscosity is low; but if you panic and try to move quickly, the viscous resistance increases considerably and you get “stuck” (Fig. 9–38). Shear thickening fluids are used in some exercise equipment—the faster you pull, the more resistance you encounter.
Derivation of the Navier–Stokes Equation for Incompressible, Isothermal Flow
From this point on, we limit our discussion to Newtonian fluids, where by definition the stress tensor is linearly proportional to the strain rate tensor. The general result (for compressible flow) is rather involved and is not included here. Instead, we assume incompressible flow (r " constant). We also assume nearly isothermal flow—namely, that local changes in temperature are small or nonexistent; this eliminates the need for a differential energy equation. A further consequence of the latter assumption is that fluid properties, such as dynamic viscosity m and kinematic viscosity n, are constant as well (Fig. 9–39). With these assumptions, it can be shown (Kundu, 1990) that the viscous stress tensor reduces to For a fluid flow that is both incompressible and isothermal: • r = constant • m = constant And therefore: • n = constant
FIGURE 9–39 The incompressible flow approximation implies constant density, and the isothermal approximation implies constant viscosity.
Viscous stress tensor for an incompressible Newtonian fluid with constant properties: tij " 2me ij
(9–55)
where eij is the strain rate tensor defined in Chap. 4. Equation 9–55 shows that stress is linearly proportional to strain. In Cartesian coordinates, the nine components of the viscous stress tensor are listed, six of which are independent due to symmetry: #u #u #v #u #w m¢ $ ≤ m¢ $ ≤ #x #y #x #z #x txz #v #u #v #v #w tyz≥ " ¶ m¢ $ ≤ 2m m¢ $ ≤ ∂ #x #y #y #z #y tzz #w #u #w #v #w m¢ $ ≤ m¢ $ ≤ 2m #x #z #y #z #z 2m
txx txy tij " £tyx tyy tzx tzy
(9–56)
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429 CHAPTER 9
In Cartesian coordinates the stress tensor of Eq. 9–53 thus becomes #u #v #u #w #u m¢ $ ≤ m¢ $ ≤ #x #y #x #z #x &P 0 0 #v #u #v #v #w 2m m¢ $ ≤ ∂ sij " £ 0 &P 0 ≥ $ ¶ m¢ $ ≤ #x #y #y #z #y 0 0 &P #w #u #w #v #w m¢ $ ≤ m¢ $ ≤ 2m #x #z #y #z #z 2m
(9–57)
Now we substitute Eq. 9–57 into the three Cartesian components of Cauchy’s equation. Let’s consider the x-component first. Equation 9–51a becomes r
Du #P #2u # #v #u # #w #u " & $ rgx $ 2m 2 $ m ¢ $ ≤ $m ¢ $ ≤ (9–58) Dt #x #y #x #y #z #x #z #x
Notice that since pressure consists of a normal stress only, it contributes only one term to Eq. 9–58. However, since the viscous stress tensor consists of both normal and shear stresses, it contributes three terms. (This is a direct result of taking the divergence of a second-order tensor, by the way.) We note that as long as the velocity components are smooth functions of x, y, and z, the order of differentiation is irrelevant. For example, the first part of the last term in Eq. 9–58 can be rewritten as m
# #w # #w ¢ ≤"m ¢ ≤ #z #x #x #z
After some clever rearrangement of the viscous terms in Eq. 9–58, r
# #v #2u Du #P #2u # #u # #w #2u " & $ rgx $ m c 2 $ $ $ 2$ $ 2d Dt #x #x #x #x #y #y #x #z #x #z "&
#P #2u #2u #2u # #u #v #w $ rgx $ m c ¢ $ $ ≤ $ 2 $ 2 $ 2 d #x #x #x #y #z #x #y #z
The term in parentheses is zero because of the continuity equation for incompressible flow (Eq. 9–17). We also recognize the last three terms as the Laplacian of velocity component u in Cartesian coordinates (Fig. 9–40). Thus, we write the x-component of the momentum equation as (9–59a)
The Laplacian Operator Cartesian coordinates: 2 2 2 2 = ∂ 2 + ∂ 2 + ∂2 ∂y ∂x ∂z Cylindrical coordinates: ∆
In similar fashion we write the y- and z-components of the momentum equation as Dv #P r "& $ rgy $ m§ 2v Dt #y
(9–59b)
Dw #P "& $ rgz $ m§ 2w Dt #z
(9–59c)
∆
Du #P "& $ rgx $ m§ 2u r Dt #x
2
2 2 = 1 ∂ ar ∂ b + 12 ∂ 2 + ∂ 2 r ∂r r ∂r ∂u ∂z
and r
respectively. Finally, we combine the three components into one vector equation; the result is the Navier–Stokes equation for incompressible flow with constant viscosity.
FIGURE 9–40 The Laplacian operator, shown here in both Cartesian and cylindrical coordinates, appears in the viscous term of the incompressible Navier–Stokes equation.
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430 FLUID MECHANICS
Incompressible Navier–Stokes equation: →
r
FIGURE 9–41 The Navier–Stokes equation is the cornerstone of fluid mechanics.
→ → DV → " & §P $ rg $ m§ 2V Dt
(9–60)
Although we derived the components of Eq. 9–60 in Cartesian coordinates, the vector form of Eq. 9–60 is valid in any orthogonal coordinate system. This famous equation is named in honor of the French engineer Louis Marie Henri Navier (1785–1836) and the English mathematician Sir George Gabriel Stokes (1819–1903), who both developed the viscous terms, although independently of each other. The Navier–Stokes equation is the cornerstone of fluid mechanics (Fig. 9–41). It may look harmless enough, but it is an unsteady, nonlinear, secondorder, partial differential equation. If we were able to solve this equation for flows of any geometry, this book would be about half as thick. Unfortunately, analytical solutions are unobtainable except for very simple flow fields. It is not too far from the truth to say that the rest of this book is devoted to solving Eq. 9–60! In fact, many researchers have spent their entire careers trying to solve the Navier–Stokes equation. Equation 9–60 has four unknowns (three velocity components and pressure), yet it represents only three equations (three components since it is a vector equation). Obviously we need another equation to make the problem solvable. The fourth equation is the incompressible continuity equation (Eq. 9–16). Before we attempt to solve this set of differential equations, we need to choose a coordinate system and expand the equations in that coordinate system.
Continuity and Navier–Stokes Equations in Cartesian Coordinates
The continuity equation (Eq. 9–16) and the Navier–Stokes equation (Eq. 9–60) are expanded in Cartesian coordinates (x, y, z) and (u, v, w): Incompressible continuity equation: #u #v #w $ $ "0 #x #y #z
(9–61a)
x-component of the incompressible Navier–Stokes equation: #u #u #u #P #2u #2u #2u #u $ u $ v $ w ≤ " & $ rgx $ m¢ 2 $ 2 $ 2 ≤ #t #x #y #z #x #x #y #z
r¢
(9–61b)
y-component of the incompressible Navier–Stokes equation: r¢
#v #v #v #P #2v #2v #2v #v $ u $ v $ w ≤ " & $ rgy $ m¢ 2 $ 2 $ 2 ≤ #t #x #y #z #y #x #y #z
(9–61c)
z-component of the incompressible Navier–Stokes equation: #w #w #w #w #P #2w #2w #2w $u $v $ w ≤ " & $ rgz $ m¢ 2 $ 2 $ 2 ≤ #t #x #y #z #z #x #y #z
r¢
(9–61d)
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431 CHAPTER 9
Continuity and Navier–Stokes Equations in Cylindrical Coordinates
The continuity equation (Eq. 9–16) and the Navier–Stokes equation (Eq. 9–60) are expanded in cylindrical coordinates (r, u, z) and (ur, uu, uz): Incompressible continuity equation:
1 #(ru r) 1 #(u u) #(u z) $ $ "0 r #r r #u #z
(9–62a)
r-component of the incompressible Navier–Stokes equation: r¢
#u r u u #u r u 2u #u r #u r $ ur $ & $ uz ≤ r #u r #t #r #z "&
#u r u r 1 #2u r 2 #u u #2u r #P 1 # $ rgr $ m c ¢r ≤ & 2$ 2 2 & 2 $ 2d r #r #r #r r r #u r #u #z
(9–62b)
u-component of the incompressible Navier–Stokes equation: #u u #u u u u #u u u ru u #u u r¢ $ uz $ ur $ $ ≤ r #u r #t #r #z "&
#u u u u 1 #2u u 2 #u r #2u u 1 #P 1 # $ rgu $ m c ¢r ≤& 2$ 2 2 $ 2 $ 2d r #u r #r #r r r #u r #u #z
(9–62c)
z-component of the incompressible Navier–Stokes equation: #u z u u #u z #u z #u z $ ur $ $ uz b ra r #u #t #r #z "&
2 2 #P 1 # #u z 1 # uz # uz $ rgz $ mc ar b $ 2 2 $ 2d r #r #r #z r #u #z
(9–62d)
The “extra” terms on both sides of the r- and u-components of the Navier–Stokes equation (Eqs. 9–62b and 9–62c) arise because of the special nature of cylindrical coordinates. Namely, as we move in the u-direction, the → unit vector er also changes direction; thus the r- and u-components are coupled (Fig. 9–42). (This coupling effect is not present in Cartesian coordinates, and thus there are no “extra” terms in Eqs. 9–61.) For completeness, the six independent components of the viscous stress tensor are listed here in cylindrical coordinates, trr tij " ° tur tzr
tru tuu tzu 2m
1 #u r # uu a b$ d r #u #r r #u r #u z ma $ b #z #r
" ¶mcr
→
er
→
eu
→
eu
r2
→
er
u2 r1
trz tuz ¢ tzz #u r #r
y
u1 x
1 #u r # uu a b$ d r #u #r r 1 #u u u r 2ma $ b r #u r #u u 1 #u z ma $ b r #u #z
mcr
#u r #u z $ b #z #r #u u 1 #u z ma $ b∂ r #u #z #u z 2m #z ma
(9–63)
FIGURE 9–42 → → Unit vectors e r and e u in cylindrical coordinates are coupled: movement in → the u-direction causes e r to change direction, and leads to extra terms in the r- and u-components of the Navier–Stokes equation.
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432 FLUID MECHANICS Three-Dimensional Incompressible Flow Four variables or unknowns: • •
Pressure P → Three components of velocity V
Four equations of motion: •
Continuity, → → • V = 0 Three components of Navier–Stokes, → → → r DV = – P + rg→ + m 2V Dt ∆
∆
•
∆
FIGURE 9–43 A general three-dimensional but incompressible flow field with constant properties requires four equations to solve for four unknowns.
9–6
■
DIFFERENTIAL ANALYSIS OF FLUID FLOW PROBLEMS
In this section we show how to apply the differential equations of motion in both Cartesian and cylindrical coordinates. There are two types of problems for which the differential equations (continuity and Navier–Stokes) are useful: • Calculating the pressure field for a known velocity field • Calculating both the velocity and pressure fields for a flow of known geometry and known boundary conditions For simplicity, we consider only incompressible flow, eliminating calculation of r as a variable. In addition, the form of the Navier–Stokes equation derived in Section 9–5 is valid only for Newtonian fluids with constant properties (viscosity, thermal conductivity, etc.). Finally, we assume negligible temperature variations, so that T is not a variable. We are left with four variables or unknowns (pressure plus three components of velocity), and we have four differential equations (Fig. 9–43).
Calculation of the Pressure Field for a Known Velocity Field
The first set of examples involves calculation of the pressure field for a known velocity field. Since pressure does not appear in the continuity equation, we can theoretically generate a velocity field based solely on conservation of mass. However, since velocity appears in both the continuity equation and the Navier–Stokes equation, these two equations are coupled. In addition, pressure appears in all three components of the Navier–Stokes equation, and thus the velocity and pressure fields are also coupled. This intimate coupling between velocity and pressure enables us to calculate the pressure field for a known velocity field. EXAMPLE 9–13
Calculating the Pressure Field in Cartesian Coordinates
Consider the steady, two-dimensional, incompressible velocity field of Example → → → 9–9, namely, V " (u, v) " (ax $ b)i $ (&ay $ cx)j . Calculate the pressure as a function of x and y.
SOLUTION For a given velocity field, we are to calculate the pressure field. Assumptions 1 The flow is steady. 2 The fluid is incompressible with constant properties. 3 The flow is two-dimensional in the xy-plane. 4 Gravity does not act in either the x- or y-direction. Analysis First we check whether the given velocity field satisfies the twodimensional, incompressible continuity equation: #u #v #w $ $ "a&a"0 #x #y #z
F
F
F a
&a
0 (2-D)
(1)
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433 CHAPTER 9
Thus, continuity is indeed satisfied by the given velocity field. If continuity were not satisfied, we would stop our analysis—the given velocity field would not be physically possible, and we could not calculate a pressure field. Next, we consider the y-component of the Navier–Stokes equation:
#v #v #v #v #P #2v #2v #2v $u $ v $ w ≤ " & $ rgy $ m¢ 2 $ 2 $ 2 ≤ #t #x #y #z #y #x #y #z
r¢
F
0
F
F
F
F
F
F
F
0 (steady) (ax $ b)c (&ay $ cx)(&a) 0 (2-D)
0
0
0 (2-D)
The y-momentum equation reduces to
#P " r(&acx & bc & a 2y $ acx) " r(&bc & a 2y) #y
(2)
The y-momentum equation is satisfied, provided we can generate a pressure field that satisfies Eq. 2. In similar fashion, the x-momentum equation reduces to
#P " r(&a 2x & ab) #x
(3)
The x-momentum equation is also satisfied, provided we can generate a pressure field that satisfies Eq. 3. In order for a steady flow solution to exist, P cannot be a function of time. Furthermore, a physically realistic steady, incompressible flow field requires a pressure field P(x, y) that is a smooth function of x and y (there can be no sudden discontinuities in either P or a derivative of P). Mathematically, this requires that the order of differentiation (x then y versus y then x) should not matter (Fig. 9–44). We check whether this is so by cross-differentiating Eqs. 2 and 3, respectively,
#2P # #P " ¢ ≤"0 #x #y #x #y
and
#2P # #P " ¢ ≤"0 #y #x #y #x
(4)
Equation 4 shows that P is a smooth function of x and y. Thus, the given velocity field satisfies the steady, two-dimensional, incompressible Navier– Stokes equation. If at this point in the analysis, the cross-differentiation of pressure were to yield two incompatible relationships, in other words if the equation in Fig. 9–44 were not satisfied, we would conclude that the given velocity field could not satisfy the steady, two-dimensional, incompressible Navier–Stokes equation, and we would abandon our attempt to calculate a steady pressure field. To calculate P (x, y), we partially integrate Eq. 2 (with respect to y) to obtain an expression for P (x, y),
Pressure field from y-momentum: P(x, y) " r¢&bcy &
a 2y 2 ≤ $ g(x) 2
(5)
Note that we add an arbitrary function of the other variable x rather than a constant of integration since this is a partial integration. We then take the partial derivative of Eq. 5 with respect to x to obtain
#P " g)(x) " r(&a 2x & ab) #x
(6)
Cross-Differentiation, xy-Plane P(x, y) is a smooth function of x and y only if the order of differentiation does not matter: ∂2P = ∂2P ∂x ∂y ∂y ∂x
FIGURE 9–44 For a two-dimensional flow field in the xy-plane, cross-differentiation reveals whether pressure P is a smooth function.
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434 FLUID MECHANICS
where we have equated our result to Eq. 3 for consistency. We now integrate Eq. 6 to obtain the function g(x):
g(x) " r¢&
a 2x 2 & abx≤ $ C 1 2
(7)
where C 1 is an arbitrary constant of integration. Finally, we substitute Eq. 7 into Eq. 5 to obtain our final expression for P(x, y ). The result is
P(x, y) ! R ¢"
a2x2 a2y2 " " abx " bcy≤ # C1 2 2
(8)
Discussion For practice, and as a check of our algebra, you should differentiate Eq. 8 with respect to both y and x, and compare to Eqs. 2 and 3. In addition, try to obtain Eq. 8 by starting with Eq. 3 rather than Eq. 2; you should get the same answer.
Notice that the final equation (Eq. 8) for pressure in Example 9–13 contains an arbitrary constant C1. This illustrates an important point about the pressure field in an incompressible flow; namely, The velocity field in an incompressible flow is not affected by the absolute magnitude of pressure, but only by pressure differences.
→
r DV = P Dt
→
∆
→
+ rg + m
→
2V
∆
FIGURE 9–45 Since pressure appears only as a gradient in the incompressible Navier–Stokes equation, the absolute magnitude of pressure is not relevant— only pressure differences matter.
This should not be surprising if we look at the Navier–Stokes equation, where P appears only as a gradient, never by itself. Another way to explain this statement is that it is not the absolute magnitude of pressure that matters, but only pressure differences (Fig. 9–45). A direct result of the statement is that we can calculate the pressure field to within an arbitrary constant, but in order to determine that constant (C1 in Example 9–13), we must measure (or otherwise obtain) P somewhere in the flow field. In other words, we require a pressure boundary condition. We illustrate this point with an example generated using computational fluid dynamics (CFD), where the continuity and Navier–Stokes equations are solved numerically (Chap. 15). Consider downward flow of air through a channel in which there is a nonsymmetrical blockage (Fig. 9–46). (Note that the computational flow domain extends much further upstream and downstream than shown in Fig. 9–46.) We calculate two cases that are identical except for the pressure condition. In case 1 we set the gage pressure far downstream of the blockage to zero. In case 2 we set the pressure at the same location to 500 Pa gage pressure. The gage pressure at the top center of the field of view and at the bottom center of the field of view are shown in Fig. 9–46 for both cases, as generated by the two CFD solutions. You can see that the pressure field for case 2 is identical to that of case 1 except that the pressure is everywhere increased by 500 Pa. Also shown in Fig. 9–46 are a velocity vector plot and a streamline plot for each case. The results are identical, confirming our statement that the velocity field is not affected by the absolute magnitude of the pressure, but only by pressure differences. Subtracting the pressure at the bottom from that at the top, we see that 'P " 12.784 Pa for both cases. The statement about pressure differences is not true for compressible flow fields, where P is the thermodynamic pressure rather than the mechanical pressure. In such cases, P is coupled with density and temperature through
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435 CHAPTER 9
an equation of state, and the absolute magnitude of pressure is important. A compressible flow solution requires not only mass and momentum conservation equations, but also an energy conservation equation and an equation of state. We take this opportunity to comment further about the CFD results shown in Fig. 9–46. You can learn a lot about the physics of fluid flow by studying relatively simple flows like this. Notice that most of the pressure drop occurs across the throat of the channel. This is caused by flow separation downstream of the blockage; rapidly moving air cannot turn around a sharp corner, and the flow separates off the walls as it exits the opening. The streamlines indicate large recirculating regions on both sides of the channel downstream of the blockage. Pressure is low in these recirculating regions. The velocity vectors indicate an inverse bell-shaped velocity profile exiting the opening—much like an exhaust jet. Because of the nonsymmetric nature of the geometry, the jet turns to the right, and the flow reattaches to the right wall much sooner than to the left wall. The pressure increases somewhat in the region where the jet impinges on the right wall, as you might expect. Finally, notice that as the air accelerates to squeeze through the orifice, the streamlines converge (as discussed in Section 9–3). As the jet of air fans out downstream, the streamlines diverge somewhat. Notice also that the streamlines in the recirculating zones are very far apart, indicating that the velocities are relatively small there; this is verified by the velocity vector plots. Finally, we note that most CFD codes do not calculate pressure by integration of the Navier–Stokes equation as we have done in Example 9–13. Instead, some kind of pressure correction algorithm is used. Most of the commonly used algorithms work by combining the continuity and Navier–Stokes equations in such a way that pressure appears in the continuity equation. The most popular pressure correction algorithms result in a form of Poisson’s equation for the change in pressure 'P from one iteration (n) to the next (n $ 1), Poisson’s equation for 'P:
§ 2('P) " RHS(n)
(9–64)
Then, as the computer iterates toward a solution, the modified continuity equation is used to “correct” the pressure field at iteration (n $ 1) from its values at iteration (n), Correction for P:
P(n$ 1) " P(n) $ 'P
Details associated with the development of pressure correction algorithms is beyond the scope of the present text. An example for two-dimensional flows is developed in Gerhart, Gross, and Hochstein (1992). EXAMPLE 9–14
Calculating the Pressure Field in Cylindrical Coordinates
Consider the steady, two-dimensional, incompressible velocity field of Example 9–5 with function f(u, t) equal to 0. This represents a line vortex whose axis lies along the z-coordinate (Fig. 9–47). The velocity components are ur " 0 and uu " K/r, where K is a constant. Calculate the pressure as a function of r and u.
SOLUTION For a given velocity field, we are to calculate the pressure field.
P = 9.222 Pa gage
P = –3.562 Pa gage (a) P = 509.222 Pa gage
P = 496.438 Pa gage (b)
FIGURE 9–46 Filled pressure contour plot, velocity vector plot, and streamlines for downward flow of air through a channel with blockage: (a) case 1; (b) case 2—identical to case 1, except P is everywhere increased by 500 Pa. On the gray-scale contour plots, dark is low pressure and light is high pressure.
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436 FLUID MECHANICS uu K uu = r
r
0
0
0
Thus, the incompressible continuity equation is satisfied. Now we look at the u component of the Navier–Stokes equation (Eq. 9–62c):
r§
#u u #u u u u #u u u ru u #u u $ uz $ ur $ $ ¥ r #u r #t #r #z K a 2b (0) r
"&
K (0) ¢& 2b r
F
0 (steady)
F
FIGURE 9–47 Streamlines and velocity profiles for a line vortex.
1 #(ru r) 1 #(u u) #(u z) $ $ "0 r #r r #u #z
Incompressible continuity:
ur = 0
Assumptions 1 The flow is steady. 2 The fluid is incompressible with constant properties. 3 The flow is two-dimensional in the ru-plane. 4 Gravity does not act in either the r- or the u-direction. Analysis The flow field must satisfy both the continuity and the momentum equations, Eqs. 9–62. For steady, two-dimensional, incompressible flow,
0
0 (2-D)
0
0
F
K
r3
K
r3
0
F
F
#u u u u 1 #2u u 2 #u r #2u u 1 #P 1 # $ rgu $ m£ ar b& 2$ 2 2 $ 2 $ 2≥ r #u r #r #r r r #u r #u #z 0 (2-D)
The u-momentum equation reduces to
u-momentum:
#P "0 #u
(1)
Thus, the u-momentum equation is satisfied, provided we can generate an appropriate pressure field that satisfies Eq. 1. In similar fashion, the rmomentum equation (Eq. 9–62b) reduces to
r-momentum: Cross-Differentiation, ru-Plane P(r, u) is a smooth function of r and u only if the order of differentiation does not matter: ∂2P = ∂2P ∂r ∂u ∂u ∂r
#P K2 "r 3 #r r
Thus, the r-momentum equation is also satisfied, provided we can generate a pressure field that satisfies Eq. 2. In order for a steady flow solution to exist, P cannot be a function of time. Furthermore, a physically realistic steady, incompressible flow field requires a pressure field P(r, u) that is a smooth function of r and u. Mathematically, this requires that the order of differentiation (r then u versus u then r) should not matter (Fig. 9–48). We check whether this is so by cross-differentiating the pressure:
# #P # 2P " a b"0 #r #u #r #u
FIGURE 9–48 For a two-dimensional flow field in the ru-plane, cross-differentiation reveals whether pressure P is a smooth function.
(2)
and
# 2P # #P " a b"0 #u #r #u #r
(3)
Equation 3 shows that P is a smooth function of r and u. Thus, the given velocity field satisfies the steady, two-dimensional, incompressible Navier– Stokes equation. We integrate Eq. 1 with respect to u to obtain an expression for P (r, u),
Pressure field from u-momentum:
P(r, u) " 0 $ g(r)
(4)
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437 CHAPTER 9
Note that we added an arbitrary function of the other variable r, rather than a constant of integration, since this is a partial integration. We take the partial derivative of Eq. 4 with respect to r to obtain
#P K2 " g)(r) " r 3 #r r
(5)
where we have equated our result to Eq. 2 for consistency. We integrate Eq. 5 to obtain the function g (r):
1 K2 g(r) " & r 2 $ C 2 r
(6)
where C is an arbitrary constant of integration. Finally, we substitute Eq. 6 into Eq. 4 to obtain our final expression for P (x, y). The result is
1 K2 P(r, U) ! " R 2 # C 2 r
P∞
P
(7)
Thus the pressure field for a line vortex decreases like 1/r 2 as we approach the origin. (The origin itself is a singularity point.) This flow field is a simplistic model of a tornado or hurricane, and the low pressure at the center is the “eye of the storm” (Fig. 9–49). We note that this flow field is irrotational, and thus Bernoulli’s equation can be used instead to calculate the pressure. If we call the pressure P* far away from the origin (r → *), where the local velocity approaches zero, Bernoulli’s equation shows that at any distance r from the origin,
Bernoulli equation:
1 P $ rV 2 " P* 2
→
2
1 K P " P* & r 2 2 r
(8)
Equation 8 agrees with our solution (Eq. 7) from the full Navier–Stokes equation if we set constant C equal to P*. A region of rotational flow near the origin would avoid the singularity there and would yield a more physically realistic model of a tornado. Discussion For practice, try to obtain Eq. 7 by starting with Eq. 2 rather than Eq. 1; you should get the same answer.
Exact Solutions of the Continuity and Navier–Stokes Equations
The remaining example problems are exact solutions of the differential equation set consisting of the incompressible continuity and Navier–Stokes equations. As you will see, these problems are by necessity simple, so that they are solvable. Most of them assume infinite boundaries and fully developed conditions so that the advective terms on the left side of the Navier–Stokes equation disappear. In addition, they are laminar, two-dimensional, and either steady or dependent on time in a predefined manner. There are six basic steps in the procedure used to solve these problems, as listed in Fig. 9–50. Step 2 is especially critical, since the boundary conditions determine the uniqueness of the solution. Step 4 is not possible analytically except for simple problems. In step 5, enough boundary conditions must be available to solve for all the constants of integration produced in step 4. Step 6 involves verifying that all the differential equations and boundary conditions are satisfied. We advise you to
r
FIGURE 9–49 The two-dimensional line vortex is a simple approximation of a tornado; the lowest pressure is at the center of the vortex.
Step 1: Set up the problem and geometry (sketches are helpful), identifying all relevant dimensions and parameters. Step 2: List all appropriate assumptions, approximations, simplifications, and boundary conditions. Step 3: Simplify the differential equations of motion (continuity and Navier–Stokes) as much as possible. Step 4: Integrate the equations, leading to one or more constants of integration. Step 5: Apply boundary conditions to solve for the constants of integration. Step 6: Verify your results.
FIGURE 9–50 Procedure for solving the incompressible continuity and Navier–Stokes equations.
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438 FLUID MECHANICS
follow these steps, even in cases where some of the steps seem trivial, in order to learn the procedure. While the examples shown here are simple, they adequately illustrate the procedure used to solve these differential equations. In Chap. 15 we discuss how computers have enabled us to solve the Navier–Stokes equations numerically for much more complicated flows using CFD. You will see that the same techniques are used there—specification of geometry, application of boundary conditions, integration of the differential equations, etc., although the steps are not always followed in the same order.
Cylinder Oil film VP
Boundary Conditions Magnifying glass
Piston
y x
FIGURE 9–51 A piston moving at speed VP in a cylinder. A thin film of oil is sheared between the piston and the cylinder; a magnified view of the oil film is shown. The no-slip boundary condition requires that the velocity of fluid adjacent to a wall equal that of the wall.
Fluid B →
n
Since boundary conditions are so critical to a proper solution, we discuss the types of boundary conditions that are commonly encountered in fluid flow analyses. The most-used boundary condition is the no-slip condition, which states that for a fluid in contact with a solid wall, the velocity of the fluid must equal that of the wall,
→
s
VB →
VA
ts, B ts, A
Fluid A
FIGURE 9–52 At an interface between two fluids, the velocity of the two fluids must be equal. In addition, the shear stress parallel to the interface must be the same in both fluids.
→
(9–65)
In other words, as its name implies, there is no “slip” between the fluid and the wall. Fluid particles adjacent to the wall adhere to the surface of the wall and move at the same velocity as the wall. A special case of Eq. 9–65 → is for a stationary wall with V wall " 0; the fluid adjacent to a stationary wall has zero velocity. For cases in which temperature effects are also considered, the temperature of the fluid must equal that of the wall, i.e., Tfluid " Twall. You must be careful to assign the no-slip condition according to your chosen frame of reference. Consider, for example, the thin film of oil between a piston and its cylinder wall (Fig. 9–51). From a stationary frame of reference, the fluid adjacent to the →cylinder→is at rest, →and the fluid adjacent to the moving piston has velocity V fluid " V wall " VP j . From a frame of reference moving with the piston, however, the fluid adjacent to the piston → has →zero velocity, but the fluid adjacent to the cylinder has velocity V fluid → " V wall " &VP j . An exception to the no-slip condition occurs in rarefied gas flows, such as during reentry of a spaceship or in the study of motion of extremely small (submicron) particles. In such flows the air can actually slip along the wall, but these flows are beyond the scope of the present text. When two fluids (fluid A and fluid B) meet at an interface, the interface boundary conditions are Interface boundary conditions:
→
→
Vfluid " Vwall
No-slip boundary condition:
→
→
VA " VB
and
ts, A " ts, B
(9–66)
where, in addition to the condition that the velocities of the two fluids must be equal, the shear stress ts acting on a fluid particle adjacent to the interface in the direction parallel to the interface must also match between the two fluids (Fig. 9–52). Note that in the figure, ts, A is drawn on the top of the fluid particle in fluid A, while ts, B is drawn on the bottom of the fluid particle in fluid B, and we have considered the direction of shear stress carefully. Because of the sign convention on shear stress, the direction of the arrows in Fig. 9–52 is opposite (a consequence of Newton’s third law). We note that although velocity is continuous across the interface, its slope is not. Also, if temperature effects are considered, TA " TB at the interface, but there may be a discontinuity in the slope of temperature at the interface as well.
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439 CHAPTER 9
What about pressure at an interface? If surface tension effects are negligible or if the interface is nearly flat, PA " PB. If the interface is sharply curved, however, as in the meniscus of liquid rising in a capillary tube, the pressure on one side of the interface can be substantially different than that on the other side. You should recall from Chap. 2 that the pressure jump across an interface is inversely proportional to the radius of curvature of the interface, as a result of surface tension effects. A degenerate form of the interface boundary condition occurs at the free surface of a liquid, meaning that fluid A is a liquid and fluid B is a gas (usually air). We illustrate a simple case in Fig. 9–53 where fluid A is liquid water and fluid B is air. The interface is flat, surface tension effects are negligible, but the water is moving horizontally (like water flowing in a calm river). In this case, the air and water velocities must match at the surface and the shear stress acting on a water particle on the surface of the water must equal that acting on an air particle just above the surface. According to Eq. 9–66, Boundary conditions at water–air interface: u water " u air
and
ts, water " m water
#u #u b " ts, air " m air b #y water #y air
Pliquid " Pgas
and
ts, liquid # 0
(9–68)
Other boundary conditions arise depending on the problem setup. For example, we often need to define inlet boundary conditions at a boundary of a flow domain where fluid enters the domain. Likewise, we define outlet boundary conditions at an outflow. Symmetry boundary conditions are useful along an axis or plane of symmetry. For example, the appropriate symmetry boundary conditions along a horizontal plane of symmetry are illustrated in Fig. 9–54. For unsteady flow problems we also need to define initial conditions (at the starting time, usually t " 0). In Examples 9–15 through 9–19, we apply boundary conditions from Eqs. 9–65 through 9–68 where appropriate. These and other boundary conditions are discussed in much greater detail in Chap. 15 where we apply them to CFD solutions. EXAMPLE 9–15
∂u ∂y b
air
uair y
uwater
∂u ∂y b
x Fluid A—water
water
u
FIGURE 9–53 Along a horizontal free surface of water and air, the water and air velocities must be equal and the shear stresses must match. However, since mair ++ mwater, a good approximation is that the shear stress at the water surface is negligibly small.
(9–67)
A quick glance at the fluid property tables reveals that mwater is over 50 times greater than mair. In order for the shear stresses to be equal, Eq. 9–67 requires that slope (#u/#y)air be more than 50 times greater than (#u/#y)water. Thus, it is reasonable to approximate the shear stress acting at the surface of the water as negligibly small compared to shear stresses elsewhere in the water. Another way to say this is that the moving water drags air along with it with little resistance from the air; in contrast, the air doesn’t slow down the water by any significant amount. In summary, for the case of a liquid in contact with a gas, and with negligible surface tension effects, the freesurface boundary conditions are Free-surface boundary conditions:
Fluid B—air
Fully Developed Couette Flow
Consider steady, incompressible, laminar flow of a Newtonian fluid in the narrow gap between two infinite parallel plates (Fig. 9–55). The top plate is moving at speed V, and the bottom plate is stationary. The distance between these two plates is h, and gravity acts in the negative z-direction (into the page in Fig. 9–55). There is no applied pressure other than hydrostatic
P = continuous
v=0
Symmetry plane y x
∂u = 0 ∂y
u
FIGURE 9–54 Boundary conditions along a plane of symmetry are defined so as to ensure that the flow field on one side of the symmetry plane is a mirror image of that on the other side, as shown here for a horizontal symmetry plane.
V Moving plate h
Fluid: r, m y Fixed plate
x
FIGURE 9–55 Geometry of Example 9–15: viscous flow between two infinite plates; upper plate moving and lower plate stationary.
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440 FLUID MECHANICS
pressure due to gravity. This flow is called Couette flow. Calculate the velocity and pressure fields, and estimate the shear force per unit area acting on the bottom plate.
SOLUTION For a given geometry and set of boundary conditions, we are to calculate the velocity and pressure fields, and then estimate the shear force per unit area acting on the bottom plate. Assumptions 1 The plates are infinite in x and z. 2 The flow is steady, i.e., #/#t of anything is zero. 3 This is a parallel flow (we assume the y-component of velocity, v, is zero). 4 The fluid is incompressible and Newtonian with constant properties, and the flow is laminar. 5 Pressure P " constant with respect to x. In other words, there is no applied pressure gradient pushing the flow in the x-direction; the flow establishes itself due to viscous stresses caused by the moving upper plate. 6 The velocity field is purely two-dimensional, meaning here that w " 0 and #/#z of any velocity component is zero. 7 Gravity acts in the negative z-direction (into the page in Fig. 9–55). We → → express this mathematically as g " &g k , or gx " gy " 0 and gz "&g. Analysis To obtain the velocity and pressure fields, we follow the step-bystep procedure outlined in Fig. 9–50. Step 1
Set up the problem and the geometry. See Fig. 9–55.
Step 2 List assumptions and boundary conditions. We have numbered and listed seven assumptions. The boundary conditions come from imposing the no-slip condition: (1) At the bottom plate (y " 0), u " v " w " 0. (2) At the top plate (y " h), u " V, v " 0, and w " 0. V
V
Step 3 Simplify the differential equations. We start with the incompressible continuity equation in Cartesian coordinates, Eq. 9–61a,
y
h
#u #x
x
FIGURE 9–56 A fully developed region of a flow field is a region where the velocity profile does not change with downstream distance. Fully developed flows are encountered in long, straight channels and pipes. Fully developed Couette flow is shown here—the velocity profile at x2 is identical to that at x1.
#u "0 #x
→
(1)
F
x = x2
#w "0 #z
$
F
x = x1
#v #y
$
assumption 3
assumption 6
Equation 1 tells us that u is not a function of x. In other words, it doesn’t matter where we place our origin—the flow is the same at any x-location. The phrase fully developed is often used to describe this situation (Fig. 9–56). This can also be obtained directly from assumption 1, which tells us that there is nothing special about any x-location since the plates are infinite in length. Furthermore, since u is not a function of time (assumption 2) or z (assumption 6), we conclude that u is at most a function of y,
u " u(y) only
Result of continuity:
(2)
We now simplify the x-momentum equation (Eq. 9–61b) as far as possible. It is good practice to list the reason for crossing out a term, as we do here:
F
#u #u #u #P #u $ u $ v $ w b "& $ rgx #t #x #y #z #x assumption 3
$ ma
V
continuity
V
assumption 2
V
V
F
ra
assumption 6 assumption 5
assumption 7
continuity
V
V
#2u #2u #2u $ $ b 2 2 #x #y #z 2 assumption 6
→
d 2u " 0 (3) dy 2
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441 CHAPTER 9
Notice that the material acceleration (left-hand side of Eq. 3) is zero, implying that fluid particles are not accelerating at all in this flow field, neither by local (unsteady) acceleration, nor by advective acceleration. Since the advective acceleration terms make the Navier–Stokes equation nonlinear, this greatly simplifies the problem. In fact, all other terms in Eq. 3 have disappeared except for a lone viscous term, which must then itself equal zero. Also notice that we have changed from a partial derivative (#/#y) to a total derivative (d/dy) in Eq. 3 as a direct result of Eq. 2. We do not show the details here, but you can show in similar fashion that every term except the pressure term in the y-momentum equation (Eq. 9–61c) goes to zero, forcing that lone term to also be zero,
#P "0 #y
(4)
In other words, P is not a function of y. Since P is also not a function of time (assumption 2) or x (assumption 5), P is at most a function of z,
P " P(z) only
Result of y-momentum:
(5)
Finally, by assumption 6 the z-component of the Navier–Stokes equation (Eq. 9–61d) simplifies to
#P " &rg #z
→
dP " &rg dz
(6)
where we used Eq. 5 to convert from a partial derivative to a total derivative. Step 4 Solve the differential equations. Continuity and y-momentum have already been “solved,” resulting in Eqs. 2 and 5, respectively. Equation 3 (x-momentum) is integrated twice to get
u " C 1y $ C 2
(7)
where C1 and C2 are constants of integration. Equation 6 (z-momentum) is integrated once, resulting in
P " &rgz $ C 3
(8)
Step 5 Apply boundary conditions. We begin with Eq. 8. Since we have specified no boundary conditions for pressure, C3 can remain an arbitrary constant. (Recall that for incompressible flow, the absolute pressure can be specified only if P is known somewhere in the flow.) For example, if we let P " P0 at z " 0, then C3 " P0 and Eq. 8 becomes
Final solution for pressure field:
P ! P0 " Rgz
(9)
Alert readers will notice that Eq. 9 represents a simple hydrostatic pressure distribution (pressure decreasing linearly as z increases). We conclude that, at least for this problem, hydrostatic pressure acts independently of the flow. More generally, we make the following statement (see also Fig. 9–57): For incompressible flow fields without free surfaces, hydrostatic pressure does not contribute to the dynamics of the flow field. In fact, in Chap. 10 we show how hydrostatic pressure can actually be removed from the equations of motion through use of a modified pressure.
Phydrostatic
z
→
g x or y
FIGURE 9–57 For incompressible flow fields without free surfaces, hydrostatic pressure does not contribute to the dynamics of the flow field.
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442 FLUID MECHANICS
We next apply boundary conditions (1) and (2) from step 2 to obtain constants C1 and C2.
Boundary condition (1): Boundary condition (2): y u =V h
FIGURE 9–58 The linear velocity profile of Example 9–15: Couette flow between parallel plates.
P
tyx txy u(y)
C 1 " V(h
Finally, Eq. 7 becomes y x
P
→
u " C1 , h $ 0 " V
Final result for velocity field:
y
C2 " 0
and
V
h
→
u " C1 , 0 $ C2 " 0
P
dy txy
dx tyx
x
P
FIGURE 9–59 Stresses acting on a differential twodimensional rectangular fluid element whose bottom face is in contact with the bottom plate of Example 9–15.
u!V
y h
(10)
The velocity field reveals a simple linear velocity profile from u " 0 at the bottom plate to u " V at the top plate, as sketched in Fig. 9–58. Step 6 Verify the results. Using Eqs. 9 and 10, you can verify that all the differential equations and boundary conditions are satisfied. To calculate the shear force per unit area acting on the bottom plate, we consider a rectangular fluid element whose bottom face is in contact with the bottom plate (Fig 9–59). Mathematically positive viscous stresses are shown. In this case, these stresses are in the proper direction since fluid above the differential element pulls it to the right while the wall below the element pulls it to the left. From Eq. 9–56, we write out the components of the viscous stress tensor,
#u #u #v #u #w V 0 m m¢ $ ≤ m¢ $ ≤ #x #y #x #z #x h #v #v #w #v #u V tij " ¶ m¢ $ ≤ 2m m¢ $ ≤ ∂ " •m 0 #x #y #y #z #y h #w #u #w #v #w 0 0 m¢ $ ≤ m¢ $ ≤ 2m #x #z #y #z #z 2m
0 0 µ (11) 0
Since the dimensions of stress are force per unit area by definition, the force per unit area acting on the bottom face of the fluid element is equal to tyx " mV/h and acts in the negative x-direction, as sketched. The shear force per unit area on the wall is equal and opposite to this (Newton’s third law); hence, →
Shear force per unit area acting on the wall:
V→ F !M i A h
(12)
The direction of this force agrees with our intuition; namely, the fluid tries to pull the bottom wall to the right, due to viscous effects (friction). Discussion The z-component of the linear momentum equation is uncoupled from the rest of the equations; this explains why we get a hydrostatic pressure distribution in the z-direction, even though the fluid is not static, but moving. Equation 11 reveals that the viscous stress tensor is constant everywhere in the flow field, not just at the bottom wall (notice that none of the components of tij is a function of location).
You may be questioning the usefulness of the final results of Example 9–15. After all, when do we encounter two infinite parallel plates, one of which is moving? Actually there are several practical flows for which the Couette flow solution is a very good approximation. One such flow occurs inside a rotational viscometer (Fig. 9–60), an instrument used to measure
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443 CHAPTER 9
viscosity. It is constructed of two concentric circular cylinders of length L— a solid, rotating inner cylinder of radius Ri and a hollow, stationary outer cylinder of radius Ro. (L is into the page in Fig. 9–60; the z-axis is out of the page.) The gap between the two cylinders is very small and contains the fluid whose viscosity is to be measured. The magnified region of Fig. 9–60 is a nearly identical setup as that of Fig. 9–55 since the gap is small, i.e. (Ro & Ri) ++ Ro. In a viscosity measurement, the angular velocity of the inner cylinder, v, is measured, as is the applied torque, Tapplied, required to rotate the cylinder. From Example 9–15, we know that the viscous shear stress acting on a fluid element adjacent to the inner cylinder is approximately equal to t " tyx # m
vR i V "m Ro & Ri Ro & Ri
vR i a2pR iLb R i Ro & Ri
R0
v
Ri
Magnifying glass t
(9–69)
where the speed V of the moving upper plate in Fig. 9–55 is replaced by the counterclockwise speed vRi of the rotating wall of the inner cylinder. In the magnified region at the bottom of Fig. 9–60, t acts to the right on the fluid element adjacent to the inner cylinder wall; hence, the force per unit area acting on the inner cylinder at this location acts to the left with magnitude given by Eq. 9–69. The total clockwise torque acting on the inner cylinder wall due to fluid viscosity is thus equal to this shear stress times the wall area times the moment arm, Tviscous " tAR i # m
Fluid: r, m
Rotating inner cylinder Stationary outer cylinder
FIGURE 9–60 A rotational viscometer; the inner cylinder rotates at angular velocity v, and a torque Tapplied is applied, from which the viscosity of the fluid is calculated.
(9–70)
Under steady conditions, the clockwise torque Tviscous is balanced by the applied counterclockwise torque Tapplied. Equating these and solving Eq. 9–70 for the fluid viscosity yields m " Tapplied
Viscosity of the fluid:
(R o & R i) 2pvR 3i L
A similar analysis can be performed on an unloaded journal bearing in which a viscous oil flows in the small gap between the inner rotating shaft and the stationary outer housing. (When the bearing is loaded, the inner and outer cylinders cease to be concentric and a more involved analysis is required.) V
EXAMPLE 9–16
Moving plate
Couette Flow with an Applied Pressure Gradient
Consider the same geometry as in Example 9–15, but instead of pressure being constant with respect to x, let there be an applied pressure gradient in the x-direction (Fig. 9–61). Specifically, let the pressure gradient in the xdirection, #P/#x, be some constant value given by
Applied pressure gradient:
#P P2 & P1 " constant " x2 & x1 #x
(1)
where x1 and x2 are two arbitrary locations along the x-axis, and P1 and P2 are the pressures at those two locations. Everything else is the same as for Example 9–15. (a) Calculate the velocity and pressure field. (b) Plot a family of velocity profiles in dimensionless form.
Fluid: r, m
h
y P1
x1
Fixed plate ∂P = P2 – P1 x2 – x1 ∂x
P2
x
x2
FIGURE 9–61 Geometry of Example 9–16: viscous flow between two infinite plates with a constant applied pressure gradient #P/#x; the upper plate is moving and the lower plate is stationary.
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444 FLUID MECHANICS
SOLUTION We are to calculate the velocity and pressure field for the flow sketched in Fig. 9–61 and plot a family of velocity profiles in dimensionless form. Assumptions The assumptions are identical to those of Example 9–15, except assumption 5 is replaced by 5 A constant pressure gradient is applied in the x-direction such that pressure changes linearly with respect to x according to Eq. 1. Analysis (a) We follow the same procedure as in Example 9–15. Much of the algebra is identical, so to save space we discuss only the differences. Step 1
See Fig. 9–61.
Step 2
Same as Example 9–15 except for assumption 5.
Step 3 The continuity equation is simplified in the same way as in Example 9–15,
Result of continuity:
u " u(y) only
(2)
The x-momentum equation is simplified in the same manner as in Example 9–15 except that the pressure gradient term remains. The result is
Result of x-momentum:
d 2u 1 #P " dy 2 m #x
(3)
Likewise, the y-momentum and z-momentum equations simplify as
Result of y-momentum:
#P "0 #y
(4)
#P " &rg #z
(5)
and
Result of z-momentum:
We cannot convert from a partial derivative to a total derivative in Eq. 5, because P is a function of both x and z in this problem, unlike in Example 9–15 where P was a function of z only. Step 4 We integrate Eq. 3 (x-momentum) twice, noting that #P/#x is a constant,
Integration of x-momentum:
u"
1 #P 2 y $ C 1y $ C 2 2m #x
(6)
where C1 and C2 are constants of integration. Equation 5 (z-momentum) is integrated once, resulting in
CAUTION! WHEN PERFORMING A PARTIAL INTEGRATION, ADD A FUNCTION OF THE OTHER VARIABLE(S)
FIGURE 9–62 A caution about partial integration.
Integration of z-momentum:
P " &rgz $ f (x)
(7)
Note that since P is now a function of both x and z, we add a function of x instead of a constant of integration in Eq. 7. This is a partial integration with respect to z, and we must be careful when performing partial integrations (Fig. 9–62). Step 5 From Eq. 7, we see that the pressure varies hydrostatically in the z-direction, and we have specified a linear change in pressure in the xdirection. Thus the function f(x) must equal a constant plus #P/#x times x. If we set P " P0 along the line x " 0, z " 0 (the y-axis), Eq. 7 becomes
Final result for pressure field:
P ! P0 #
&P x " Rgz &x
(8)
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445 CHAPTER 9
We next apply the velocity boundary conditions (1) and (2) from step 2 of Example 9–15 to obtain constants C1 and C2.
Boundary condition (1): u"
1 #P , 0 $ C1 , 0 $ C2 " 0 2m #x
→
C2 " 0
and
Boundary condition (2): u"
1 #P 2 h $ C1 , h $ 0 " V 2m #x
→
C1 "
V 1 #P & h h 2m #x
Finally, Eq. 6 becomes
u!
Vy 1 &P 2 # ( y " hy) h 2M &x
V (9) u(y)
Equation 9 indicates that the velocity field consists of the superposition of two parts: a linear velocity profile from u " 0 at the bottom plate to u " V at the top plate, and a parabolic distribution that depends on the magnitude of the applied pressure gradient. If the pressure gradient is zero, the parabolic portion of Eq. 9 disappears and the profile is linear, just as in Example 9–15; this is sketched as the dashed line in Fig. 9–63. If the pressure gradient is negative (pressure decreasing in the x-direction, causing flow to be pushed from left to right), #P/#x + 0 and the velocity profile looks like the one sketched in Fig. 9–63. A special case is when V " 0 (top plate stationary); the linear portion of Eq. 9 vanishes, and the velocity profile is parabolic and symmetric about the center of the channel (y " h/2); this is sketched as the dotted line in Fig. 9–63. Step 6 You can use Eqs. 8 and 9 to verify that all the differential equations and boundary conditions are satisfied. (b) We use dimensional analysis to generate the dimensionless groups (groups). We set up the problem in terms of velocity component u as a function of y, h, V, m, and #P/#x. There are six variables (including the dependent variable u), and since there are three primary dimensions represented in the problem (mass, length, and time), we expect 6 & 3 " 3 dimensionless groups. When we pick h, V, and m as our repeating variables, we get the following result using the method of repeating variables (details are left to the reader—this is a good review of Chap. 7 material):
Result of dimensional analysis:
y h 2 #P u "fa , b V h mV #x
(10)
Using these three dimensionless groups, we rewrite Eq. 9 as
Dimensionless form of velocity field:
1 u* ! y* # P*y*(y* " 1) 2
(11)
where the dimensionless parameters are
u* "
u V
y* "
y h
P* "
h 2 #P mV #x
In Fig. 9–64, u* is plotted as a function of y* for several values of P*, using Eq. 11.
h y x
FIGURE 9–63 The velocity profile of Example 9–16: Couette flow between parallel plates with an applied negative pressure gradient; the dashed line indicates the profile for a zero pressure gradient, and the dotted line indicates the profile for a negative pressure gradient with the upper plate stationary (V " 0).
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446 FLUID MECHANICS 1
0.8 P* = 15 0.6
–15 10
y* = y/h
–10
5 0
0.4
FIGURE 9–64 Nondimensional velocity profiles for Couette flow with an applied pressure gradient; profiles are shown for several values of nondimensional pressure gradient.
0.2
0 –1.5
u(y) h
y x
FIGURE 9–65 The velocity profile for fully developed two-dimensional channel flow (planar Poiseuille flow).
–1
z x P = Patm
Fixed wall
Air
→
g
h
FIGURE 9–66 Geometry of Example 9–17: a viscous film of oil falling by gravity along a vertical wall.
–0.5
0
0.5 u* = u/V
1
1.5
2
2.5
Discussion When the result is nondimensionalized, we see that Eq. 11 represents a family of velocity profiles. We also see that when the pressure gradient is positive (flow being pushed from right to left) and of sufficient magnitude, we can have reverse flow in the bottom portion of the channel. For all cases, the boundary conditions reduce to u* ! 0 at y* ! 0 and u* ! 1 at y* ! 1. If there is a pressure gradient but both walls are stationary, the flow is called two-dimensional channel flow, or planar Poiseuille flow (Fig. 9–65). We note, however, that most authors reserve the name Poiseuille flow for fully developed pipe flow—the axisymmetric analog of two-dimensional channel flow (see Example 9–18).
EXAMPLE 9–17
Oil film: r, m
–5
Oil Film Flowing Down a Vertical Wall by Gravity
Consider steady, incompressible, parallel, laminar flow of a film of oil falling slowly down an infinite vertical wall (Fig. 9–66). The oil film thickness is h, and gravity acts in the negative z-direction (downward in Fig. 9–66). There is no applied (forced) pressure driving the flow—the oil falls by gravity alone. Calculate the velocity and pressure fields in the oil film and sketch the normalized velocity profile. You may neglect changes in the hydrostatic pressure of the surrounding air.
SOLUTION For a given geometry and set of boundary conditions, we are to calculate the velocity and pressure fields and plot the velocity profile. Assumptions 1 The wall is infinite in the yz-plane (y is into the page for a right-handed coordinate system). 2 The flow is steady (all partial derivatives with respect to time are zero). 3 The flow is parallel (the x-component of velocity, u, is zero everywhere). 4 The fluid is incompressible and Newtonian with constant properties, and the flow is laminar. 5 Pressure P ! Patm ! constant at the free surface. In other words, there is no applied pressure gradient pushing the flow; the flow establishes itself due to a balance between gravitational forces and viscous forces. In addition, since there is no gravity force in the horizontal direction, P ! Patm everywhere. 6 The velocity field is purely
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447 CHAPTER 9
two-dimensional, which implies that velocity component v " 0 and all partial derivatives with respect to y are zero. 7 Gravity acts in the negative z-direction. → → We express this mathematically as g " &g k , or gx " gy " 0 and gz " &g. Analysis We obtain the velocity and pressure fields by following the step-bystep procedure for differential fluid flow solutions. (Fig. 9–50). Step 1
Set up the problem and the geometry. See Fig. 9–66.
Step 2 List assumptions and boundary conditions. We have listed seven assumptions. The boundary conditions are: (1) There is no slip at the wall; at x " 0, u " v " w " 0. (2) At the free surface (x " h), there is negligible shear (Eq. 9–68), which for a vertical free surface in this coordinate system means #w/#x " 0 at x " h. Step 3 Write out and simplify the differential equations. We start with the incompressible continuity equation in Cartesian coordinates,
#u #x
#v #y
$
$
#w "0 #z
→
#w "0 #z
(1)
F
F assumption 3
assumption 6
Equation 1 tells us that w is not a function of z; i.e., it doesn’t matter where we place our origin—the flow is the same at any z-location. In other words, the flow is fully developed. Since w is not a function of time (assumption 2), z (Eq. 1), or y (assumption 6), we conclude that w is at most a function of x,
w " w(x) only
Result of continuity:
(2)
We now simplify each component of the Navier–Stokes equation as far as possible. Since u " v " 0 everywhere, and gravity does not act in the x- or y-directions, the x- and y-momentum equations are satisfied exactly (in fact all terms are zero in both equations). The z-momentum equation reduces to
#w #w #w #P #w $ u $ v $ w b" & $ rgz #t #x #y #z #z
ra
F
assumption 6
continuity
assumption 5
→
&rg
d 2w rg " m dx 2
(3)
V
V
#2w #2w #2w $ ma 2 $ $ b 2 #x #y #z 2
NOTICE
V
assumption 3
V
V
V
F
assumption 2
assumption 6
continuity
The material acceleration (left side of Eq. 3) is zero, implying that fluid particles are not accelerating in this flow field, neither by local nor advective acceleration. Since the advective acceleration terms make the Navier–Stokes equation nonlinear, this greatly simplifies the problem. We have changed from a partial derivative (#/#x) to a total derivative (d/dx) in Eq. 3 as a direct result of Eq. 2, reducing the partial differential equation (PDE) to an ordinary differential equation (ODE). ODEs are of course much easier than PDEs to solve (Fig. 9–67). Step 4 Solve the differential equations. The continuity and x- and y-momentum equations have already been “solved.” Equation 3 (z-momentum) is integrated twice to get
w"
rg 2 x $ C 1x $ C 2 2m
(4)
If u = u(x) only, change from PDE to ODE: ∂u ∂x
du dx
FIGURE 9–67 In Examples 9–15 through 9–18, the equations of motion are reduced from partial differential equations to ordinary differential equations, making them much easier to solve.
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448 FLUID MECHANICS 0
Step 5 Apply boundary conditions. We apply boundary conditions (1) and (2) from step 2 to obtain constants C1 and C2,
–0.1
Boundary condition (1):
C2 " 0
and
–0.2 w*
w " 0 $ 0 $ C2 " 0
Boundary condition (2): –0.3
Finally, Eq. 4 becomes –0.4 –0.5 –0.6 0.2
0.4
0.6
→
rgh C1 " & m
Rgx Rg 2 Rg hx ! x " (x " 2h) M 2M 2M
(5)
Since x + h in the film, w is negative everywhere, as expected (flow is downward). The pressure field is trivial; namely, P ! Patm everywhere.
Free surface
0
w!
Velocity field:
Wall
rg dw h $ C1 " 0 b " m dx x"h
0.8
1
x*
FIGURE 9–68 The normalized velocity profile of Example 9–17: an oil film falling down a vertical wall.
Step 6 Verify the results. You can verify that all the differential equations and boundary conditions are satisfied. We normalize Eq. 5 by inspection: we let x* " x/h and w* " wm/(rgh2). Equation 5 becomes
Normalized velocity profile:
w* "
x* (x* & 2) 2
(6)
We plot the normalized velocity field in Fig. 9–68. Discussion The velocity profile has a large slope near the wall due to the no-slip condition there (w " 0 at x " 0), but zero slope at the free surface, where the boundary condition is zero shear stress (#w/#x " 0 at x " h). We could have introduced a factor of &2 in the definition of w* so that w* would equal 1 instead of &12 at the free surface.
The solution procedure used in Examples 9–15 through 9–17 in Cartesian coordinates can also be used in any other coordinate system. In Example 9–18 we present the classic problem of fully developed flow in a round pipe, for which we use cylindrical coordinates. Pipe wall
EXAMPLE 9–18
Fluid: r, m D
r x
V P1 x1
R ∂P ∂x
=
P2 – P1
P2 x2
x2 – x1
FIGURE 9–69 Geometry of Example 9–18: steady laminar flow in a long round pipe with an applied pressure gradient #P/#x pushing fluid through the pipe. The pressure gradient is usually caused by a pump and/or gravity.
Fully Developed Flow in a Round Pipe— Poiseuille Flow
Consider steady, incompressible, laminar flow of a Newtonian fluid in an infinitely long round pipe of diameter D or radius R " D/2 (Fig. 9–69). We ignore the effects of gravity. A constant pressure gradient #P/#x is applied in the x-direction,
Applied pressure gradient:
#P P2 & P1 " constant " x2 & x1 #x
(1)
where x1 and x2 are two arbitrary locations along the x-axis, and P1 and P2 are the pressures at those two locations. Note that we adopt a modified cylindrical coordinate system here with x instead of z for the axial component, namely, (r, u, x) and (ur , uu, u). Derive an expression for the velocity field inside the pipe and estimate the viscous shear force per unit surface area acting on the pipe wall.
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449 CHAPTER 9
SOLUTION For flow inside a round pipe we are to calculate the velocity field, and then estimate the viscous shear stress acting on the pipe wall. Assumptions 1 The pipe is infinitely long in the x-direction. 2 The flow is steady (all partial time derivatives are zero). 3 This is a parallel flow (the r-component of velocity, ur , is zero). 4 The fluid is incompressible and Newtonian with constant properties, and the flow is laminar. 5 A constant-pressure gradient is applied in the x-direction such that pressure changes linearly with respect to x according to Eq. 1. 6 The velocity field is axisymmetric with no swirl, implying that uu " 0 and all partial derivatives with respect to u are zero. 7 We ignore the effects of gravity. Analysis To obtain the velocity field, we follow the step-by-step procedure outlined in Fig. 9–50. Step 1
Lay out the problem and the geometry. See Fig. 9–69.
Step 2 List assumptions and boundary conditions. We have listed seven assumptions. The first boundary condition comes from imposing the no-slip → condition at the pipe wall: (1) at r " R, V " 0. The second boundary condition comes from the fact that the centerline of the pipe is an axis of symmetry: (2) at r " 0, du/dr " 0. Step 3 Write out and simplify the differential equations. We start with the incompressible continuity equation in cylindrical coordinates, a modified version of Eq. 9–62a,
1 #(ru r) 1 #(u u) #u $ $ "0 r #r r #u #x
assumption 3
assumption 6
→
#u "0 #x
(2)
Equation 2 tells us that u is not a function of x. In other words, it doesn’t matter where we place our origin—the flow is the same at any x-location. This can also be inferred directly from assumption 1, which tells us that there is nothing special about any x-location since the pipe is infinite in length—the flow is fully developed. Furthermore, since u is not a function of time (assumption 2) or u (assumption 6), we conclude that u is at most a function of r,
u " u(r) only
Result of continuity:
(3)
We now simplify the axial momentum equation (a modified version of Eq. 9–62d) as far as possible:
assumption 2 assumption 3
V
V
V
u u #u #u #u #u $ ur $ $ u b r #u #t #r #x
F
ra
assumption 6
"&
continuity
F
#P 1 # #u 1 #2u #2u $ rgx $ ma ar b $ 2 2 $ 2b r #r #x #r r #u #x
F
assumption 7
assumption 6 continuity
or
1 d du 1 #P ar b " m #x r dr dr
(4)
As in Examples 9–15 through 9–17, the material acceleration (entire left side of the x-momentum equation) is zero, implying that fluid particles are
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450 FLUID MECHANICS The Navier–Stokes Equation + (V A ∂V ∂t
→ → →
)V
∆
•
B= –
→
→
∆
r
P + rg + m
∆
→
→
2V
Nonlinear term
FIGURE 9–70 For incompressible flow solutions in which the advective terms in the Navier–Stokes equation are zero, the equation becomes linear since the advective term is the only nonlinear term in the equation.
not accelerating at all in this flow field, and linearizing the Navier–Stokes equation (Fig. 9–70). We have replaced the partial derivative operators for the u-derivatives with total derivative operators because of Eq. 3. In similar fashion, every term in the r-momentum equation (Eq. 9–62b) except the pressure gradient term is zero, forcing that lone term to also be zero,
#P "0 #r
r-momentum:
(5)
In other words, P is not a function of r. Since P is also not a function of time (assumption 2) or u (assumption 6), P can be at most a function of x,
P " P(x) only
Result of r-momentum:
(6)
Therefore, we can replace the partial derivative operator for the pressure gradient in Eq. 4 by the total derivative operator since P varies only with x. Finally, all terms of the u-component of the Navier–Stokes equation (Eq. 9–62c) go to zero. Step 4 Solve the differential equations. Continuity and r-momentum have already been “solved,” resulting in Eqs. 3 and 6, respectively. The u-momentum equation has vanished, and thus we are left with Eq. 4 (x-momentum). After multiplying both sides by r, we integrate once to obtain
r
du r 2 dP " $ C1 dr 2m dx
(7)
where C1 is a constant of integration. Note that the pressure gradient dP/dx is a constant here. Dividing both sides of Eq. 7 by r, we integrate a second time to get
u"
r 2 dP $ C 1 ln r $ C 2 4m dx
(8)
where C2 is a second constant of integration. Step 5 Apply boundary conditions. First, we apply boundary condition (2) to Eq. 7,
Boundary condition (2):
umax
Boundary condition (1):
r x u(r)
→
C1 " 0
An alternative way to interpret this boundary condition is that u must remain finite at the centerline of the pipe. This is possible only if constant C1 is equal to 0, since ln(0) is undefined in Eq. 8. Now we apply boundary condition (1),
V = uavg = umax/2
D
0 " 0 $ C1
u
R2 dP $ 0 $ C2 " 0 4m dx
→
R2 dP C2 " & 4m dx
Finally, Eq. 7 becomes
R
Axial velocity:
FIGURE 9–71 Axial velocity profile of Example 9–18: steady laminar flow in a long round pipe with an applied constantpressure gradient dP/dx pushing fluid through the pipe.
u"
u!
1 dP 2 (r " R 2) 4M dx
(9)
The axial velocity profile is thus in the shape of a paraboloid, as sketched in Fig. 9–71. Step 6 Verify the results. You can verify that all the differential equations and boundary conditions are satisfied.
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451 CHAPTER 9
We calculate some other properties of fully developed laminar pipe flow as well. For example, the maximum axial velocity obviously occurs at the centerline of the pipe (Fig. 9–71). Setting r " 0 in Eq. 9 yields
u max " &
Maximum axial velocity:
R2 dP 4m dx
(10)
The volume flow rate through the pipe is found by integrating Eq. 9 through the whole cross-sectional area of the pipe,
# V"
2p
! ! u"0
R
ur dr du "
r"0
2p dP 4m dx
!
R
(r 2 & R2)r dr " &
r"0
pR4 dP 8m dx
(11)
Since volume flow rate is also equal to the average axial velocity times crosssectional area, we can easily determine the average axial velocity V:
Average axial velocity:
# V (&pR4(8m) (dP(dx) R2 dP V" " "& 2 A 8m dx pR
(12)
Comparing Eqs. 10 and 12 we see that for fully developed laminar pipe flow, the average axial velocity is equal to exactly half of the maximum axial velocity. To calculate the viscous shear force per unit surface area acting on the pipe wall, we consider a differential fluid element adjacent to the bottom portion of the pipe wall (Fig. 9–72). Pressure stresses and mathematically positive viscous stresses are shown. From Eq. 9–63 (modified for our coordinate system), we write the viscous stress tensor as
trr tij " £tur txr
tru tuu txu
0
0
#u #r 0 µ
0
0
0
trx tux ≥ " • 0 txx #u m #r
m
trx " m
du R dP " dr 2 dx
(13)
(14)
For flow from left to right, dP/dx is negative, so the viscous shear stress on the bottom of the fluid element at the wall is in the direction opposite to that indicated in Fig. 9–72. (This agrees with our intuition since the pipe wall exerts a retarding force on the fluid.) The shear force per unit area on the wall is equal and opposite to this; hence, →
Viscous shear force per unit area acting on the wall:
R dP → F !" i A 2 dx
trx
(15)
The direction of this force again agrees with our intuition; namely, the fluid tries to pull the bottom wall to the right, due to friction, when dP/dx is negative. Discussion Since du/dr " 0 at the centerline of the pipe, trx " 0 there. You are encouraged to try to obtain Eq. 15 by using a control volume approach instead, taking your control volume as the fluid in the pipe between any two
P
r
P + dP dx dx 2
txr
dr
P – dP dx dx 2
txr
dx x
We use Eq. 9 for u, and set r " R at the pipe wall; component trx of Eq. 13 reduces to
Viscous shear stress at the pipe wall:
Centerline
P
trx
Pipe wall
FIGURE 9–72 Pressure and viscous shear stresses acting on a differential fluid element whose bottom face is in contact with the pipe wall.
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452 FLUID MECHANICS Pipe wall CV
Fluid: r, m
x-locations, x1 and x2 (Fig. 9–73). You should get the same answer. (Hint: Since the flow is fully developed, the axial velocity profile at location 1 is identical to that at location 2.) Note that when the volume flow rate through the pipe exceeds a critical value, instabilities in the flow occur, and the solution presented here is no longer valid. Specifically, flow in the pipe becomes turbulent rather than laminar; turbulent pipe flow is discussed in more detail in Chap. 8. This problem is also solved in Chap. 8 using an alternative approach.
r x
P1 x1
R dP dx
=
P2
P2 – P1
x2
x2 – x1
FIGURE 9–73 Control volume used to obtain Eq. 15 of Example 9–18 by an alternative method.
So far, all our Navier–Stokes solutions have been for steady flow. You can imagine how much more complicated the solutions must get if the flow is allowed to be unsteady, and the time derivative term in the Navier–Stokes equation does not disappear. Nevertheless, there are some unsteady flow problems that can be solved analytically. We present one of these in Example 9–19. EXAMPLE 9–19
→
→
g = –gk
Fluid: r, m V
z
x
Infinite flat plate
FIGURE 9–74 Geometry and setup for Example 9–19; the y-coordinate is into the page.
Sudden Motion of an Infinite Flat Plate
Consider a viscous Newtonian fluid on top of an infinite flat plate lying in the xy-plane at z " 0 (Fig. 9–74). The fluid is at rest until time t " 0, when the plate suddenly starts moving at speed V in the x-direction. Gravity acts in the &z-direction. Determine the pressure and velocity fields.
SOLUTION The velocity and pressure fields are to be calculated for the case of fluid on top of an infinite flat plate that suddenly starts moving. Assumptions 1 The wall is infinite in the x- and y-directions; thus, nothing is special about any particular x- or y-location. 2 The flow is parallel everywhere (w " 0). 3 Pressure P " constant with respect to x. In other words, there is no applied pressure gradient pushing the flow in the x-direction; flow occurs due to viscous stresses caused by the moving plate. 4 The fluid is incompressible and Newtonian with constant properties, and the flow is laminar. 5 The velocity field is two-dimensional in the xz-plane; therefore, v " 0, and all partial derivatives with respect to y are zero. 6 Gravity acts in the &z-direction. Analysis To obtain the velocity and pressure fields, we follow the step-bystep procedure outlined in Fig. 9–50. Step 1
Lay out the problem and the geometry. (See Fig. 9–74.)
Step 2 List assumptions and boundary conditions. We have listed six assumptions. The boundary conditions are: (1) At t " 0, u " 0 everywhere (no flow until the plate starts moving); (2) at z " 0, u " V for all values of x and y (no-slip condition at the plate); (3) as z → *, u " 0 (far from the plate, the effect of the moving plate is not felt); and (4) at z " 0, P " Pwall (the pressure at the wall is constant at any x- or y-location along the plate). Step 3 Write out and simplify the differential equations. We start with the incompressible continuity equation in Cartesian coordinates (Eq. 9–61a),
#u #v $ #x #y
$
#w "0 #z
→
#u "0 #x
(1)
F
F
assumption 5
assumption 2
Equation 1 tells us that u is not a function of x. Furthermore, since u is not a function of y (assumption 5), we conclude that u is at most a function of z and t,
cen72367_ch09.qxd 11/4/04 7:17 PM Page 453
453 CHAPTER 9
u " u (z, t) only
Result of continuity:
(2)
The y-momentum equation reduces to
#P "0 #y
(3)
by assumptions 5 and 6 (all terms with v, the y-component of velocity, vanish, and gravity does not act in the y-direction). Equation 3 simply tells us that pressure is not a function of y; hence,
P " P(z, t) only
Result of y-momentum:
(4)
Similarly the z-momentum equation reduces to
#P " &rg #z
(5)
We now simplify the x-momentum equation (Eq. 9–61b) as far as possible.
#u #u #u #P #u $ u $ v $ w b" & $ rgx #t #x #y #z #x
ra
assumption 2
Equatio
assumption 3 assumption 6
→
r
#2u #u "m 2 #t #z
(6)
V
V
#2u #2u #2u $ ma 2 $ $ b #x #y 2 #z 2 continuity
assumption 5
It is convenient to combine the viscosity and density into the kinematic viscosity, defined as n " m/r. Equation 6 reduces to the well-known onedimensional diffusion equation (Fig. 9–75),
#2u #u "n 2 #t #z
Result of x-momentum:
(8)
where we have added a function of time instead of a constant of integration since P is a function of two variables, z and t (see Eq. 4). Equation 7 (x-momentum) is a linear partial differential equation whose solution is obtained by combining the two independent variables z and t into one independent variable. The result is called a similarity solution, the details of which are beyond the scope of this text. Note that the one-dimensional diffusion equation occurs in many other fields of engineering, such as diffusion of species (mass diffusion) and diffusion of heat (conduction); details about the solution can be found in books on these subjects. The solution of Eq. 7 is intimately tied to the boundary condition that the plate is impulsively started, and the result is
Integration of x-momentum:
z u " C 1 c1 & erfa bd 22nt
(9)
where erf in Eq. 9 is the error function (Çengel, 2003), defined as
Error function:
erf(j) "
! 2p 2
0
j 2
e &h dh
fusion
Dif The 1-D
n
Equatio
2u ∂u = n ∂ 2 ∂z ∂t
(7)
Step 4 Solve the differential equations. Continuity and y-momentum have already been “solved,” resulting in Eqs. 2 and 4, respectively. Equation 5 (z-momentum) is integrated once, resulting in
P " &rgz $ f (t)
e Day
n of th
F
assumption 5
V
V
V
V
continuity
(10)
FIGURE 9–75 The one-dimensional diffusion equation is linear, but it is a partial differential equation (PDE). It occurs in many fields of science and engineering.
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454 FLUID MECHANICS 1
The error function is commonly used in probability theory and is plotted in Fig. 9–76. Tables of the error function can be found in many reference books, and some calculators and spreadsheets can calculate the error function directly. It is also provided as a function in the EES software that comes with this text.
0.8
0.6
Step 5 Apply boundary conditions. We begin with Eq. 8 for pressure. Boundary condition (4) requires that P " Pwall at z " 0 for all times, and Eq. 8 becomes
erf(j) 0.4
0.2
→
P " 0 $ f (t) " Pwall
Boundary condition (4):
f (t) " Pwall
In other words, the arbitrary function of time, f(t), turns out not to be a function of time at all, but merely a constant. Thus,
P ! Pwall " Rgz
Final result for pressure field:
0 0
0.5
1
1.5 j
2
2.5
3
FIGURE 9–76 The error function ranges from 0 at j " 0 to 1 as j → *.
(11)
which is simply hydrostatic pressure. We conclude that hydrostatic pressure acts independently of the flow. Boundary conditions (1) and (3) from step 2 have already been applied in order to obtain the solution of the x-momentum equation in step 4. Since erf(0) " 0, the second boundary condition yields
Boundary condition (2):
u " C 1(1 & 0) " V
→
C1 " V
and Eq. 9 becomes
Final result for velocity field:
24 h 0.15 8h 3h
Normalized variables:
1h 0.05
Normalized velocity field:
30 s 0 0.2
0.4 0.6 u, m/s
u* "
u V
and
z* "
z 22nt
Then we rewrite Eq. 12 in terms of nondimensional parameters:
15 min 5 min
0
(12)
Several velocity profiles are plotted in Fig. 9–77 for the specific case of water at room temperature (n " 1.004 , 10&6 m2/s) with V " 1.0 m/s. At t " 0, there is no flow. As time goes on, the motion of the plate is felt farther and farther into the fluid, as expected. Notice how long it takes for viscous diffusion to penetrate into the fluid—after 15 min of flow, the effect of the moving plate is not felt beyond about 10 cm above the plate! We define normalized variables u* and z* as
0.2
z, m 0.1
z u ! V c1 " erf a bd 22Nt
0.8
FIGURE 9–77 Velocity profiles of Example 9–19: flow of water above an impulsively started infinite plate; n " 1.004 , 10&6 m2/s and V " 1.0 m/s.
1
u* ! 1 " erf (z*)
(13)
The combination of unity minus the error function occurs often in engineering and is given the special name complementary error function and symbol erfc. Thus Eq. 13 can also be written as
Alternative form of the velocity field:
u* " erfc (z*)
(14)
The beauty of the normalization is that this one equation for u* as a function of z* is valid for any fluid (with kinematic viscosity n) above a plate moving at any speed V and at any location z in the fluid at any time t! The normalized velocity profile of Eq. 13 is sketched in Fig. 9–78. All the profiles of Fig. 9–77 collapse into the single profile of Fig. 9–78; such a profile is called a similarity profile. Step 6 Verify the results. You can verify that all the differential equations and boundary conditions are satisfied.
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455 CHAPTER 9
Discussion The time required for momentum to diffuse into the fluid seems much longer than we would expect based on our intuition. This is because the solution presented here is valid only for laminar flow. It turns out that if the plate’s speed is large enough, or if there are significant vibrations in the plate or disturbances in the fluid, the flow will become turbulent. In a turbulent flow, large eddies mix rapidly moving fluid near the wall with slowly moving fluid away from the wall. This mixing process occurs rather quickly, so that turbulent diffusion is usually orders of magnitude faster than laminar diffusion.
3 2.5 2 z 1.5 2 √ nt 1 0.5
Examples 9–15 through 9–19 are for incompressible laminar flow. The same set of differential equations (incompressible continuity and Navier– Stokes) is valid for incompressible turbulent flow. However, turbulent flow solutions are much more complicated because the flow contains random, unsteady, three-dimensional eddies that mix the fluid. Furthermore, these eddies may range in size over several orders of magnitude. In a turbulent flow field, none of the terms in the equations can be ignored (with the exception of the gravity term in some cases), and thus our only hope of obtaining a solution is through numerical computations on a computer. Computational fluid dynamics (CFD) is discussed in Chap. 15.
0 0
0.2
0.4
0.6
0.8
1
u/V
FIGURE 9–78 Normalized velocity profile of Example 9–19: laminar flow of a viscous fluid above an impulsively started infinite plate.
SUMMARY In this chapter we derive the differential forms of conservation of mass (the continuity equation) and conservation of linear momentum (the Navier–Stokes equation). For incompressible flow of a Newtonian fluid with constant properties, the continuity equation is →
→
§%V"0 and the Navier–Stokes equation is →
r
→ → DV → " &§P $ rg $ m§ 2V Dt
For incompressible two-dimensional flow, we also define the stream function c. In Cartesian coordinates, u"
#c #y
v"&
#c #x
We show that the difference in the value of c from one streamline to another is equal to the volume flow rate per unit
width between the two streamlines and that curves of constant c are streamlines of the flow. We provide several examples showing how the differential equations of fluid motion are used to generate an expression for the pressure field for a given velocity field and to generate expressions for both velocity and pressure fields for a flow with specified geometry and boundary conditions. The solution procedure learned here can be extended to much more complicated flows whose solutions require the aid of a computer. The Navier–Stokes equation is the cornerstone of fluid mechanics. Although we have the necessary differential equations that describe fluid flow (continuity and Navier–Stokes), it is another matter to solve them. For some simple (usually infinite) geometries, the equations reduce to equations that we can solve analytically. For more complicated geometries, the equations are nonlinear, coupled, second-order, partial differential equations that cannot be solved with pencil and paper. We must then resort to either approximate solutions (Chap. 10) or numerical solutions (Chap. 15).
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456 FLUID MECHANICS
REFERENCES AND SUGGESTED READING 1. R. W. Fox and A. T. McDonald. Introduction to Fluid Mechanics, 5th ed. New York: Wiley, 1998.
5. P. K. Kundu. Fluid Mechanics. San Diego, CA: Academic Press, 1990.
2. P. M. Gerhart, R. J. Gross, and J. I. Hochstein. Fundamentals of Fluid Mechanics, 2nd ed. Reading, MA: Addison-Wesley, 1992.
6. R. L. Panton. Incompressible Flow, 2nd ed. New York: Wiley, 1996.
3. R. J. Heinsohn and J. M. Cimbala. Indoor Air Quality Engineering. New York: Marcel-Dekker, 2003.
7. M. R. Spiegel. Vector Analysis, Schaum’s Outline Series, Theory and Problems. New York: McGraw-Hill Trade, 1968.
4. Y. A. Çengel. Heat Transfer: A Practical Approach, 2nd ed. New York: McGraw-Hill, 2003.
8. M. Van Dyke. An Album of Fluid Motion. Stanford, CA: The Parabolic Press, 1982.
PROBLEMS* General and Mathematical Background Problems 9–1C Explain the fundamental differences between a flow domain and a control volume. 9–2C What does it mean when we say that two or more differential equations are coupled? 9–3C For a three-dimensional, unsteady, incompressible flow field in which temperature variations are insignificant, how many unknowns are there? List the equations required to solve for these unknowns. 9–4C For a three-dimensional, unsteady, compressible flow field in which temperature and density variations are significant, how many unknowns are there? List the equations required to solve for these unknowns. (Hint: Assume other flow properties like viscosity and thermal conductivity can be treated as constants.) 9–5C
The divergence theorem is expressed as
! § % G dV " " G % n dA →
→
→
→
V
→
A
where G is a vector, V is a volume, and A is the surface area that encloses and defines the volume. Express the divergence theorem in words.
9–7 A Taylor series expansion of function f (x) about some x-location x0 is given as df f(x0 $ dx) " f(x0) $ a b dx dx x"x0 $
1 d 2f 1 d 3f a 2b dx 2 $ a 3b dx 3 $ p 2! dx x"x0 3! dx x"x0
Consider the function f(x) " exp(x) " ex. Suppose we know the value of f(x) at x " x0, i.e., we know the value of f (x0), and we want to estimate the value of this function at some x location near x0. Generate the first four terms of the Taylor series expansion for the given function (up to order dx3 as in the above equation). For x0 " 0 and dx " &0.1, use your truncated Taylor series expansion to estimate f(x0 $ dx). Compare your result with the exact value of e&0.1. How many digits of accuracy do you achieve with your truncated Taylor series? →
→
→
→
→
→
→
* Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
→
9–9 Let vector G be given by G " 4xzi & y 2 j $ yzk and let V be the volume of a cube of unit length with its corner at
→
9–6 Transform the position x " (4, 3, &4) from Cartesian (x, y, z) coordinates to cylindrical (r, u, z) coordinates, → including units. The values of x are in units of meters.
→
→
by G " 2xzi & 12 x 2 j $ z 2 k . 9–8 Let vector G be given → Calculate the divergence of G , and simplify as much as possible. Is there anything special about your result? Answer: 0
A
V
1
y x z
1 1
FIGURE P9–9
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457 CHAPTER 9
the origin, bounded by x " 0 to 1, y " 0 to 1, and z " 0 to 1 (Fig. P9–9). Area A is the surface area of the cube. Perform both integrals of the divergence theorem and verify that they are equal. Show all your work. 9–10 The product rule can →be applied to→the →divergence of → → → → scalar f times vector G as: ! % ( fG ) " G % !f $ f! % G . Expand both sides of this equation in Cartesian coordinates and verify that it is correct. 9–11 The outer product of two vectors is a second-order tensor with nine components. In Cartesian coordinates, it is FxG x F G " CFyG x FzG x
→→
FxG y FyG y FzG y
FxG z FyG zS FzG z
9–16 A steady, two-dimensional, incompressible velocity field has Cartesian velocity components u " Cy/(x2 $ y2) and v " &Cx/(x2 $ y2), where C is a constant. Transform these Cartesian velocity components into cylindrical velocity components ur and uu, simplifying as much as possible. You should recognize this flow. What kind of flow is this? Answer: 0, &C/r, line vortex
The product rule applied to the divergence→ of the product→ of → → →→ → two→ vectors F →and→ G can be written as ! % ( F G ) " G (! → % F ) $ (F % !)G . Expand both sides of this equation in Cartesian coordinates and verify that it is correct. 9–12 Use the product rule→of →Prob. 9–11 to show that → →→ →→ → → ! % (rV V ) " V ! % (rV ) $ r(V % !)V . 9–13 On many occasions we need to transform a velocity from Cartesian (x, y, z) coordinates to cylindrical (r, u, z) coordinates (or vice versa). Using Fig. P9–13 as a guide, transform cylindrical velocity components (ur, uu, uz) into Cartesian velocity components (u, v, w). (Hint: Since the z-component of velocity remains the same in such a transformation, we need only to consider the xy-plane, as in Fig. P9–13.)
y
9–15 Beth is studying a rotating flow in a wind tunnel. She measures the u and v components of velocity using a hot-wire anemometer. At x " 0.50 m and y " 0.20 m, u " 10.3 m/s and v " &5.6 m/s. Unfortunately, the data analysis program requires input in cylindrical coordinates (r, u) and (ur, uu). Help Beth transform her data into cylindrical coordinates. Specifically, calculate r, u, ur, and uu at the given data point.
9–17 Consider a spiraling line vortex/sink flow in the xy- or ru-plane as sketched in Fig. P9–17. The two-dimensional cylindrical velocity components (ur, uu) for this flow field are ur " C/2pr and uu " ./2pr, where C and . are constants (m is negative and . is positive). Transform these two-dimensional cylindrical velocity components into two-dimensional Cartesian velocity components (u, v). Your final answer should contain no r or u—only x and y. As a check of your algebra, calculate V 2 using Cartesian coordinates, and compare to V 2 obtained from the given velocity components in cylindrical components. y
v
x →
V uu
ur r
FIGURE P9–17
u
u x
FIGURE P9–13 9–14 Using Fig. P9–13 as a guide, transform Cartesian velocity components (u, v, w) into cylindrical velocity components (ur, uu, uz). (Hint: Since the z-component of velocity remains the same in such a transformation, we need only to consider the xy-plane.)
9–18E Alex is measuring the time-averaged velocity components in a pump using a laser Doppler velocimeter (LDV). Since the laser beams are aligned with the radial and tangential directions of the pump, he measures the ur and uu components of velocity. At r " 6.20 in and u " 30.0°, ur " 1.37 ft/s and uu " 3.82 ft/s. Unfortunately, the data analysis program requires input in Cartesian coordinates (x, y) in feet and (u, v) in ft/s. Help Alex transform his data into Cartesian coordinates. Specifically, calculate x, y, u, and v at the given data point.
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458 FLUID MECHANICS
Continuity Equation
Time = t1
9–19C If a flow field is compressible, what can we say about the material derivative of density? What about if the flow field is incompressible?
Time = t3
9–20C In this chapter we derive the continuity equation in two ways: by using the divergence theorem and by summing mass flow rates through each face of an infinitesimal control volume. Explain why the former is so much less involved than the latter.
Time = t2
FIGURE P9–24
9–21
A two-dimensional diverging duct is being designed to diffuse the high-speed air exiting a wind tunnel. The x-axis is the centerline of the duct (it is symmetric about the x-axis), and the top and bottom walls are to be curved in such a way that the axial wind speed u decreases approximately linearly from u1 " 300 m/s at section 1 to u2 " 100 m/s at section 2 (Fig. P9–21). Meanwhile, the air density r is to increase approximately linearly from r1 " 0.85 kg/m3 at section 1 to r2 " 1.2 kg/m3 at section 2. The diverging duct is 2.0 m long and is 1.60 m high at section 1 (only the upper half is sketched in Fig. P9–21; the half-height at section 1 is 0.80 m). (a) Predict the y-component of velocity, v(x, y), in the duct. (b) Plot the approximate shape of the duct, ignoring friction on the walls. (c) What should be the half-height of the duct at section 2?
∆x = 2.0 m
9–25 Verify that the spiraling line vortex/sink flow in the ru-plane of Prob. 9–17 satisfies the two-dimensional incompressible continuity equation. What happens to conservation of mass at the origin? Discuss. 9–26 Verify that the steady, two-dimensional, incompressible velocity field of Prob. 9–16 satisfies the continuity equation. Stay in Cartesian coordinates and show all your work. 9–27 Consider the steady, two-dimensional velocity field → → → given by V " (u, v) " (1.3 $ 2.8x)i $ (1.5 & 2.8y)j . Verify that this flow field is incompressible. 9–28 Imagine a steady, two-dimensional, incompressible flow that is purely radial in the xy- or ru-plane. In other words, velocity component ur is nonzero, but uu is zero everywhere (Fig. P9–28). What is the most general form of velocity component ur that does not violate conservation of mass? y
0.8 m y (1)
ur x
r
(2)
x
FIGURE P9–21
9–22 Repeat Example 9–1 (gas compressed in a cylinder by a piston), but without using the continuity equation. Instead, consider the fundamental definition of density as mass divided by volume. Verify that Eq. 5 of Example 9–1 is correct. 9–23 The→ compressible form of the continuity equation is → (#r/#t) $ ! % (rV ) " 0. Expand this equation as far as possible in Cartesian coordinates (x, y, z) and (u, v, w). 9–24 In Example 9–6 we derive the equation for volumetric → → strain rate, (1/V)(DV/Dt) " ! % V . Write this as a word equation and discuss what happens to the volume of a fluid element as it moves around in a compressible fluid flow field (Fig. P9–24).
u
FIGURE P9–28 9–29 Consider the following steady, three-dimensional veloc-→ → " (u, v, w) " (axy2 & b)i ity field in Cartesian coordinates: V → → 3 $ cy j $ dxyk , where a, b, c, and d are constants. Under what conditions is this flow field incompressible? Answer: a " &3c
9–30 The u velocity component of a steady, two-dimensional, incompressible flow field is u " ax $ b, where a and b are constants. Velocity component v is unknown. Generate an expression for v as a function of x and y.
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459 CHAPTER 9
9–31 Imagine a steady, two-dimensional, incompressible flow that is purely circular in the xy- or ru-plane. In other words, velocity component uu is nonzero, but ur is zero everywhere (Fig. P9–31). What is the most general form of velocity component uu that does not violate conservation of mass? y
uu
r u x
FIGURE P9–31 9–32 The u velocity component of a steady, two-dimensional, incompressible flow field is u ! ax # by, where a and b are constants. Velocity component v is unknown. Generate an expression for v as a function of x and y. Answer: "ay # f (x)
9–33 The u velocity component of a steady, two-dimensional, incompressible flow field is u ! ax2 " bxy, where a and b are constants. Velocity component v is unknown. Generate an expression for v as a function of x and y. 9–34 Consider steady flow of water through an axisymmetric garden hose nozzle (Fig. P9–34). The axial component of velocity increases linearly from uz, entrance to uz, exit as sketched. Between z ! 0 and z ! L, the axial velocity component is given by uz ! uz, entrance # [(uz, exit " uz, entrance)/L]z. Generate an expression for the radial velocity component ur between z ! 0 and z ! L. You may ignore frictional effects on the walls. Dentrance
z=0
9–36C What is significant about curves of constant stream function? Explain why the stream function is useful in fluid mechanics. 9–37C What restrictions or conditions are imposed on stream function c so that it exactly satisfies the two-dimensional incompressible continuity equation by definition? Why are these restrictions necessary? 9–38C Consider two-dimensional flow in the xy-plane. What is the significance of the difference in value of stream function c from one streamline to another? 9–39 There are numerous occasions in which a fairly uniform free-stream flow of speed V in the x-direction encounters a long circular cylinder of radius a aligned normal to the flow (Fig. P9–39). Examples include air flowing around a car antenna, wind blowing against a flag pole or telephone pole, wind hitting electric wires, and ocean currents impinging on the submerged round beams that support oil platforms. In all these cases, the flow at the rear of the cylinder is separated and unsteady and usually turbulent. However, the flow in the front half of the cylinder is much more steady and predictable. In fact, except for a very thin boundary layer near the cylinder surface, the flow field can be approximated by the following steady, two-dimensional stream function in the xy- or ru-plane, with the cylinder centered at the origin: c ! V sin u(r " a2/r). Generate expressions for the radial and tangential velocity components. y V
r u x r=a
FIGURE P9–39 Dexit
r z uz, entrance
Stream Function
uz, exit
9–40 Consider fully developed Couette flow—flow between two infinite parallel plates separated by distance h, with the top plate moving and the bottom plate stationary as illustrated in Fig. P9–40. The flow is steady, incompressible, and twodimensional in the xy-plane. The velocity field is given by V
z=L
FIGURE P9–34 9–35 Two velocity components of a steady, incompressible flow field are known: u ! ax # bxy # cy2 and v ! axz " byz2, where a, b, and c are constants. Velocity component w is missing. Generate an expression for w as a function of x, y, and z.
h
u =V
y h
y x
FIGURE P9–40
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460 FLUID MECHANICS →
→
→
V " (u, v) " (Vy/h)i $ 0j . Generate an expression for stream function c along the vertical dashed line in Fig. P9–40. For convenience, let c " 0 along the bottom wall of the channel. What is the value of c along the top wall? Answers: Vy 2/2h, Vh/2
9–41 As a follow-up to Prob. 9–40, calculate the volume flow rate per unit width into the page of Fig. P9–40 from first principles (integration of the velocity field). Compare your result to that obtained directly from the stream function. Discuss. 9–42E Consider the Couette flow of Fig. P9–40. For the case in which V " 10.0 ft/s and h " 1.20 in, plot several streamlines using evenly spaced values of stream function. Are the streamlines themselves equally spaced? Discuss why or why not. 9–43 Consider fully developed, two-dimensional channel flow—flow between two infinite parallel plates separated by distance h, with both the top plate and bottom plate stationary, and a forced pressure gradient dP/dx driving the flow as illustrated in Fig. P9–43. (dP/dx is constant and negative.) The flow is steady, incompressible, and two-dimensional in the xy-plane. The velocity components are given by u " (1/2m)(dP/dx)(y2 & hy) and v " 0, where m is the fluid’s viscosity. Generate an expression for stream function c along the vertical dashed line in Fig. P9–43. For convenience, let c " 0 along the bottom wall of the channel. What is the value of c along the top wall?
Answers: 0.00234 m3/s, 10.0 m/s Dividing streamlines c = cu
Sampling probe
h
Vfree stream
Vavg
⋅ V
c = cl
FIGURE P9–46 9–47 Suppose the suction applied to the sampling probe of Prob. 9–46 were too weak instead of too strong. Sketch what the streamlines would look like in that case. What would you call this kind of sampling? Label the lower and upper dividing streamlines.
u(y) h
air speed through the probe should be the same as that of the airstream (isokinetic sampling). However, if the applied suction is too large, as sketched in Fig. P9–46, the air speed through the probe is greater than that of the airstream (superisokinetic sampling). For simplicity consider a twodimensional case in which the sampling probe height is h " 4.5 mm and its width (into the page of Fig. P9–46) is W " 52 mm. The values of the stream function corresponding to the lower and upper dividing streamlines are cl " 0.105 m2/s and cu " 0.150 m2/s, respectively. Calculate the volume flow rate through the probe (in units of m3/s) and the average speed of the air sucked through the probe.
y x
FIGURE P9–43 9–44 As a follow-up to Prob. 9–43, calculate the volume flow rate per unit width into the page of Fig. P9–43 from first principles (integration of the velocity field). Compare your result to that obtained directly from the stream function. Discuss. 9–45 Consider the channel flow of Fig. P9–43. The fluid is water at 20/C. For the case in which dP/dx " &20,000 N/m3 and h " 1.20 mm, plot several streamlines using evenly spaced values of stream function. Are the streamlines themselves equally spaced? Discuss why or why not. 9–46 In the field of air pollution control, one often needs to sample the quality of a moving airstream. In such measurements a sampling probe is aligned with the flow as sketched in Fig. P9–46. A suction pump draws air through the probe at # volume flow rate V as sketched. For accurate sampling, the
9–48 Consider the air sampling probe of Prob. 9–46. If the upper and lower streamlines are 5.8 mm apart in the airstream far upstream of the probe, estimate the free stream speed Vfree stream. 9–49 Consider a steady, two-dimensional, incompressible flow field called a uniform stream. The fluid speed is V everywhere, and the flow is aligned with the x-axis (Fig. P9–49). The Cartesian velocity components are u " V and y c2
V
c1 c0 = 0 –c1 –c2
FIGURE P9–49
x
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461 CHAPTER 9
v " 0. Generate an expression for the stream function for this flow. Suppose V " 8.9 m/s. If c2 is a horizontal line at y " 0.5 m and the value of c along the x-axis is zero, calculate the volume flow rate per unit width (into the page of Fig. P9–49) between these two streamlines. 9–50E Consider the steady, two-dimensional, incompressible flow field of Prob. 9–33, for which the u velocity component is u " ax2 & bxy, where a " 0.45 (ft · s)&1, and b " 0.75 (ft · s)&1. Let v " 0 for all values of x when y " 0 (that is, v " 0 along the x-axis). Generate an expression for the stream function and plot some streamlines of the flow. For consistency, set c " 0 along the x-axis, and plot in the range 0 + x + 3 ft and 0 + y + 4 ft. 9–51 A uniform stream of speed V is inclined at angle a from the x-axis (Fig. P9–51). The flow is steady, two-dimensional, and incompressible. The Cartesian velocity components are u " V cos a and v " V sin a. Generate an expression for the stream function for this flow.
c2 V
63.7%
c = 4.15
h
c = 2.80
c = 2.03
FIGURE P9–56 9–57 If the average velocity in the main branch of the duct of Prob. 9–56 is 11.4 m/s, calculate duct height h in units of cm. Obtain your result in two ways, showing all your work. You may use the results of Prob. 9–56 in only one of the methods.
y
c1
9–56 A steady, incompressible, two-dimensional CFD calculation of flow through an asymmetric two-dimensional branching duct reveals the streamline pattern sketched in Fig. P9–56, where the values of c are in units of m2/s, and W is the width of the duct into the page. The values of stream function c on the duct walls are shown. What percentage of the flow goes through the upper branch of the duct? Answer:
a x
c0 –c1 –c2
FIGURE P9–51 9–52 A steady, two-dimensional, incompressible flow field in the xy-plane has the following stream function: c " ax2 $ bxy $ cy2, where a, b, and c are constants. (a) Obtain expressions for velocity components u and v. (b) Verify that the flow field satisfies the incompressible continuity equation.
9–58 Consider steady, incompressible, axisymmetric flow (r, z) and (ur, uz) for which the stream function is defined as ur " &(1/r)(#c/#z) and uz " (1/r)(#c/#r). Verify that c so defined satisfies the continuity equation. What conditions or restrictions are required on c? 9–59 Consider steady, incompressible, two-dimensional flow due to a line source at the origin (Fig. P9–59). Fluid is created at the origin and spreads out radially in all directions in the xy-plane.# The net volume flow rate of created fluid per unit width is V (L (into the page of Fig. P9–59), where L is the width of the line source into the page in Fig. P9–59. y
9–53
For the velocity field of Prob. 9–52, plot streamlines c " 0, 1, 2, 3, 4, 5, and 6 m2/s. Let constants a, b, and c have the following values: a " 0.50 s&1, b " &1.3 s&1, and c " 0.50 s&1. For consistency, plot streamlines between x " &2 and 2 m, and y " &4 and 4 m. Indicate the direction of flow with arrows.
⋅ V L
r u
9–54 A steady, two-dimensional, incompressible flow field in the xy-plane has a stream function given by c " ax2 & by2 $ cx $ dxy, where a, b, c, and d are constants. (a) Obtain expressions for velocity components u and v. (b) Verify that the flow field satisfies the incompressible continuity equation. 9–55 Repeat Prob. 9–54, except make up your own stream function. You may create any function c(x, y) that you desire, as long as it contains at least three terms and is not the same as an example or problem in this text. Discuss.
ur
x
FIGURE P9–59
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462 FLUID MECHANICS
Since mass must be conserved everywhere except at the origin (a singularity point), the volume flow rate# per unit width through a circle of any radius r must also be V (L. If we (arbitrarily) specify stream function c to be zero along the positive x-axis (u " 0), what is the value of c along the positive y-axis (u " 90/)? What is the value of c along the negative xaxis (u " 180/)? 9–60 Repeat Prob. 9–59 for the case of a line sink instead # of a line source. Let V (L be a positive value, but the flow is everywhere in the opposite direction.
9–65 Streaklines are shown in Fig. P9–65 for flow of water over the front portion of a blunt, axisymmetric cylinder aligned with the flow. Streaklines are generated by introducing air bubbles at evenly spaced points upstream of the field of view. Only the top half is shown since the flow is symmetric about the horizontal axis. Since the flow is steady, the streaklines are coincident with streamlines. Discuss how you can tell from the streamline pattern whether the flow speed in a particular region of the flow field is (relatively) large or small.
9–61 Consider the garden hose nozzle of Prob. 9–34. Generate an expression for the stream function corresponding to this flow field. 9–62E
Consider the garden hose nozzle of Probs. 9–34 and 9–61. Let the entrance and exit nozzle diameters be 0.50 and 0.14 in, respectively, and let the nozzle length be 2.0 in. The volume flow rate through the nozzle is 2.0 gal/min. (a) Calculate the axial speeds (ft/s) at the nozzle entrance and at the nozzle exit. (b) Plot several streamlines in the rz-plane inside the nozzle, and design the appropriate nozzle shape. 9–63 Flow separates at a sharp corner along a wall and forms a recirculating separation bubble as sketched in Fig. P9–63 (streamlines are shown). The value of the stream function at the wall is zero, and that of the uppermost streamline shown is some positive value cupper. Discuss the value of the stream function inside the separation bubble. In particular, is it positive or negative? Why? Where in the flow is c a minimum?
c = cupper
FIGURE P9–65 Courtesy ONERA. Photograph by Werlé.
9–66E A sketch of flow streamlines (contours of constant stream function) is shown in Fig. P9–66E for steady, incompressible, two-dimensional flow of air in a curved duct. (a) Draw arrows on the streamlines to indicate the direction of flow. (b) If h " 2.0 in, what is the approximate speed of the air at point P? (c) Repeat part (b) if the fluid were water instead of air. Discuss. Answers: (b) 0.78 ft/s, (c) 0.78 ft/s
c = 0.32 ft2/s
P h c=0 Separation bubble
c = 0.45 ft2/s
FIGURE P9–66E
FIGURE P9–63 9–64 A graduate student is running a CFD code for his MS research project and generates a plot of flow streamlines (contours of constant stream function). The contours are of equally spaced values of stream function. Professor I. C. Flows looks at the plot and immediately points to a region of the flow and says, “Look how fast the flow is moving here!” What did Professor Flows notice about the streamlines in that region and how did she know that the flow was fast in that region?
9–67 We briefly mention the compressible stream function cr in this chapter, defined in Cartesian coordinates as ru " (#cr /#y) and rv " &(#cr /#x). What are the primary dimensions of cr? Write the units of cr in primary SI units and in primary English units. 9–68 In Example 9–2 we provide expressions for u, v, and r for flow through a compressible converging duct. Generate an expression for the compressible stream function cr that describes this flow field. For consistency, set cr " 0 along the x-axis.
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463 CHAPTER 9
9–69
In Prob. 9–21 we developed expressions for u, v, and r for flow through the compressible, twodimensional, diverging duct of a high-speed wind tunnel. Generate an expression for the compressible stream function cr that describes this flow field. For consistency, set cr " 0 along the x-axis. Plot several streamlines and verify that they agree with those you plotted in Prob. 9–21. What is the value of cr at the top wall of the diverging duct?
fluid is air at room temperature. Note that contours of constant compressible stream function are plotted in Fig. P9–71, even though the flow itself is approximated as incompressible. Values of cr are in units of kg/m · s. (a) Draw an arrow on the plot to indicate the direction and relative magnitude of the velocity at point A. Repeat for point B. (b) What is the approximate speed of the air at point B? (Point B is between streamlines 5 and 6 in Fig. P9–71.)
9–70 Steady, incompressible, two-dimensional flow over a newly designed small hydrofoil of chord length c " 9.0 mm is modeled with a commercial computational fluid dynamics (CFD) code. A close-up view of flow streamlines (contours of constant stream function) is shown in Fig. P9–70. Values of the stream function are in units of m2/s. The fluid is water at room temperature. (a) Draw an arrow on the plot to indicate the direction and relative magnitude of the velocity at point A. Repeat for point B. Discuss how your results can be used to explain how such a body creates lift. (b) What is the approximate speed of the air at point A? (Point A is between streamlines 1.65 and 1.66 in Fig. P9–70.)
Linear Momentum Equations, Boundary Conditions, and Applications
A
1.70
1.71
1.68
1.69
1.66
1.67 c
1.64
1.65 1.63
1.62 B
1.60
9–72C The general control volume equation for conservation of linear momentum is
!
→
rg dV $
CV
!
→
sij % n dA
CS
I
II
"
!
CV
→ # arVb dV $ #t III
!
CS
→ →
→
arVb V % n dA
IV
Discuss the meaning of each term in this equation. The terms are labeled for convenience. Write the equation as a word equation. 9–73C An airplane flies through the sky at constant veloc→ ity V airplane (Fig. P9–73C). Discuss the velocity boundary conditions on the air adjacent to the surface of the airplane from two frames of reference: (a) standing on the ground, and (b) moving with the airplane. Likewise, what are the far-field velocity boundary conditions of the air (far away from the airplane) in both frames of reference?
1.61 →
Vairplane
FIGURE P9–70 9–71 Time-averaged, turbulent, incompressible, two-dimensional flow over a square block of dimension h " 1 m is modeled with a commercial computational fluid dynamics (CFD) code. A close-up view of flow streamlines (contours of constant stream function) is shown in Fig. P9–71. The
FIGURE P9–73C 9–74C What are constitutive equations, and to which fluid mechanics equation are they applied? 9–75C What is mechanical pressure Pm, and how is it used in an incompressible flow solution? 9–76C What is the main distinction between a Newtonian fluid and a non-Newtonian fluid? Name at least three Newtonian fluids and three non-Newtonian fluids.
3 h
h
2
FIGURE P9–71
10 6 4 A
9–77C Define or describe each type of fluid: (a) viscoelastic fluid, (b) pseudoplastic fluid, (c) dilatant fluid, (d) Bingham plastic fluid.
B 1
9–78 A stirrer mixes liquid chemicals in a large tank (Fig. P9–78). The free surface of the liquid is exposed to room air. Surface tension effects are negligible. Discuss the boundary conditions required to solve this problem. Specifically, what
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464 FLUID MECHANICS
are the velocity boundary conditions in terms of cylindrical coordinates (r, u, z) and velocity components (ur, uu, uz) at all surfaces, including the blades and the free surface? What pressure boundary conditions are appropriate for this flow field? Write mathematical equations for each boundary condition and discuss. P = Patm
Free surface
Rtank D z
#u u 1 #u r tru " tur " mc a & u ub $ d r #u #r
(2)
Are these the same? In other words is Eq. 2 equivalent to Eq. 1, or do these other authors define their viscous stress tensor differently? Show all your work. 9–82 Engine oil at T " 60/C is forced between two very large, stationary, parallel flat plates separated by a thin gap height h " 2.5 mm (Fig. P9–82). The plate dimensions are L " 1.5 m and W " 0.75 m. The outlet pressure is atmospheric, and the inlet pressure is 1 atm gage pressure. Estimate the volume flow rate of oil. Also calculate the Reynolds number of the oil flow, based on gap height h and average velocity V. Is the flow laminar or turbulent? Answers: 9.10
v
r, m
Some authors write this component instead as
r
FIGURE P9–78
, 10&4 m3/s, 14.5, laminar
9–79 Repeat Prob. 9–78, but from a frame of reference rotating with the stirrer blades at angular velocity v. 9–80 Consider liquid in a cylindrical tank. Both the tank and the liquid rotate as a rigid body (Fig. P9–80). The free surface of the liquid is exposed to room air. Surface tension effects are negligible. Discuss the boundary conditions required to solve this problem. Specifically, what are the velocity boundary conditions in terms of cylindrical coordinates (r, u, z) and velocity components (ur, uu, uz) at all surfaces, including the tank walls and the free surface? What pressure boundary conditions are appropriate for this flow field? Write mathematical equations for each boundary condition and discuss.
y
W L x
V
FIGURE P9–82 9–83 Consider the following steady, two-dimensional, incom→ → " (u, v) " (ax $ b)i $ (&ay $ pressible velocity field: V → cx2)j , where a, b, and c are constants. Calculate the pressure as a function of x and y. Answer: cannot be found 9–84 Consider the following steady, two-dimensional, incom→ → → pressible velocity field: V " (u, v) " (&ax2)i $ (2axy)j , where a is a constant. Calculate the pressure as a function of x and y.
v
Free surface
Pout
h
Pin
9–85 Consider steady, two-dimensional, incompressible flow due to a spiraling line vortex/sink flow centered on the
P = Patm
uu
→
g
r uu =
R Liquid
K r
z r
r
FIGURE P9–80 9–81 The ru-component of the viscous stress tensor in cylindrical coordinates is given by tru " tur " m cr
1 #u r # uu a b$ d r #u #r r
(1)
ur =
C r
FIGURE P9–85
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465 CHAPTER 9
z-axis. Streamlines and velocity vectors are shown in Fig. P9–85. The velocity field is ur " C/r and uu " K/r, where C and K are constants. Calculate the pressure as a function of r and u.
Oil film: r, m
9–86 Consider the steady, two-dimensional, incompressible → → → velocity field, V " (u, v) " (ax $ b)i $ (&ay $ c)j , where a, b, and c are constants. Calculate the pressure as a function of x and y.
Air
n
→
g
9–87 Consider steady, incompressible, parallel, laminar flow of a viscous fluid falling between two infinite vertical walls (Fig. P9–87). The distance between the walls is h, and gravity acts in the negative z-direction (downward in the figure). There is no applied (forced) pressure driving the flow— the fluid falls by gravity alone. The pressure is constant everywhere in the flow field. Calculate the velocity field and sketch the velocity profile using appropriate nondimensionalized variables.
P = Patm
Fixed wall
s
h
z
a
x
FIGURE P9–89 9–90 For the falling oil film of Prob. 9–89, generate an expression for the volume flow rate per unit width of oil # falling down# the wall (V (L) as a function of r, m, h, and g. Calculate (V (L) for an oil film of thickness 5.0 mm with r " 888 kg/m3 and m " 0.80 kg/m · s.
z x
9–91 Fixed wall
Fluid: r, m
Fixed wall →
g
h
The first two viscous terms in the u-component of the #u u uu 1 # ar b & 2d. Navier–Stokes equation (Eq. 9–62c) are m c r #r #r r
Expand this expression as far as possible using the product rule, yielding three terms. Now combine all three terms into one term. (Hint: Use the product rule in reverse—some trial and error may be required.) 9–92 An incompressible Newtonian liquid is confined between two concentric circular cylinders of infinite length—
FIGURE P9–87 Liquid: r, m
9–88 For the fluid falling between two parallel vertical walls (Prob. 9–87), generate an expression for the volume # flow rate per unit width (V (L) as a function of r, m, h, and g. Compare your result to that of the same fluid falling along one vertical wall with a free surface replacing the second wall (Example 9–17), all else being equal. Discuss the differences and provide a physical explanation. Answer: rgh3/12m
Ro
vi
Ri
downward
9–89 Repeat Example 9–17, except for the case in which the wall is inclined at angle a (Fig. P9–89). Generate expressions for both the pressure and velocity fields. As a check, make sure that your result agrees with that of Example 9–17 when a " 90/. [Hint: It is most convenient to use the (s, y, n) coordinate system with velocity components (us, v, un), where y is into the page in Fig. P9–89. Plot the dimensionless velocity profile u*s versus n* for the case in which a " 60/.]
Rotating inner cylinder Stationary outer cylinder
FIGURE P9–92
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466 FLUID MECHANICS
a solid inner cylinder of radius Ri and a hollow, stationary outer cylinder of radius Ro (Fig. P9–92; the z-axis is out of the page). The inner cylinder rotates at angular velocity vi. The flow is steady, laminar, and two-dimensional in the ruplane. The flow is also rotationally symmetric, meaning that nothing is a function of coordinate u (uu and P are functions of radius r only). The flow is also circular, meaning that velocity component ur " 0 everywhere. Generate an exact expression for velocity component uu as a function of radius r and the other parameters in the problem. You may ignore gravity. (Hint: The result of Prob. 9–91 is useful.) 9–93 Analyze and discuss two limiting cases of Prob. 9–92: (a) The gap is very small. Show that the velocity profile approaches linear from the outer cylinder wall to the inner cylinder wall. In other words, for a very tiny gap the velocity profile reduces to that of simple two-dimensional Couette flow. (Hint: Define y " Ro & r, h " gap thickness " Ro & Ri, and V " speed of the “upper plate” " Ri vi.) (b) The outer cylinder radius becomes infinite, while the inner cylinder radius becomes very small. What kind of flow does this approach? 9–94 Repeat Prob. 9–92 for the more general case. Namely, let the inner cylinder rotate at angular velocity vi and let the outer cylinder rotate at angular velocity vo. All else is the same as Prob. 9–92. Generate an exact expression for velocity component uu as a function of radius r and the other parameters in the problem. Verify that when vo " 0 your result simplifies to that of Prob. 9–92. 9–95 Analyze and discuss a limiting case of Prob. 9–94 in which there is no inner cylinder (Ri " vi " 0). Generate an expression for uu as a function of r. What kind of flow is this? Describe how this flow could be set up experimentally. Answer: vor
9–96 Consider steady, incompressible, laminar flow of a Newtonian fluid in an infinitely long round pipe annulus of inner radius Ri and outer radius Ro (Fig. P9–96). Ignore the effects of gravity. A constant negative pressure gradient #P/#x is applied in the x-direction, (#P/dx) " (P2 & P1)/(x2 & x1), where x1 and x2 are two arbitrary locations along the x-axis,
and P1 and P2 are the pressures at those two locations. The pressure gradient may be caused by a pump and/or gravity. Note that we adopt a modified cylindrical coordinate system here with x instead of z for the axial component, namely, (r, u, x) and (ur, uu, u). Derive an expression for the velocity field in the annular space in the pipe. 9–97 Consider again the pipe annulus sketched in Fig. P9–96. Assume that the pressure is constant everywhere (there is no forced pressure gradient driving the flow). However, let the inner wall be moving at steady velocity V to the right. The outer wall is still stationary. (This is a kind of axisymmetric Couette flow.) Generate an expression for the x-component of velocity u as a function of r and the other parameters in the problem. 9–98 Repeat Prob. 9–97 except swap the stationary and moving walls. In particular, let the inner wall be stationary, and let the outer pipe wall be moving at steady velocity V to the right, all else being equal. Generate an expression for the x-component of velocity u as a function of r and the other parameters in the problem. 9–99 Consider a modified form of Couette flow in which there are two immiscible fluids sandwiched between two infinitely long and wide, parallel flat plates (Fig. P9–99). The flow is steady, incompressible, parallel, and laminar. The top plate moves at velocity V to the right, and the bottom plate is stationary. Gravity acts in the &z-direction (downward in the figure). There is no forced pressure gradient pushing the fluids through the channel—the flow is set up solely by viscous effects created by the moving upper plate. You may ignore surface tension effects and assume that the interface is horizontal. The pressure at the bottom of the flow (z " 0) is equal to P0. (a) List all the appropriate boundary conditions on both velocity and pressure. (Hint: There are six required boundary conditions.) (b) Solve for the velocity field. (Hint: Split up the solution into two portions, one for each fluid. Generate expressions for u1 as a function of z and u2 as a function of z.) (c) Solve for the pressure field. (Hint: Again split up the solution. Solve for P1 and P2.) (d) Let fluid 1 be water and let fluid 2 be unused engine oil, both at 80/C. Also let h1 " 5.0 mm, h2 " 8.0 mm, and V " 10.0 m/s. Plot u as a function of z across the entire channel. Discuss the results.
Outer pipe wall Fluid: r, m
Moving wall
r x
P1 x1
FIGURE P9–96
Ri
Ro
∂P P2 – P1 = ∂x x2 – x1
Interface
h2
V Fluid 2
r2, m2
Fluid 1
r1, m1
P2 x2
h2
z
x
FIGURE P9–99
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467 CHAPTER 9
9–100 Consider steady, incompressible, laminar flow of a Newtonian fluid in an infinitely long round pipe of diameter D or radius R ! D/2 inclined at angle a (Fig. P9–100). There is no applied pressure gradient ("P/"x ! 0). Instead, the fluid flows down the pipe due to gravity alone. We adopt the coordinate system shown, with x down the axis of the pipe. Derive an expression for the x-component of velocity u as a function of radius r and the other parameters of the problem. Calculate the volume flow rate and average axial velocity through the pipe. Answers: rg (sin
a)(R2
#
r 2)/4m,
rg (sin
a)pR4/8m,
rg (sin
a)R2/8m
Pipe wall Fluid: r, m
D
r R →
g
x
a
FIGURE P9–100
Review Problems 9–101C Explain why the incompressible flow approximation and the constant temperature approximation usually go hand in hand. 9–102C For each part, write the official name for the differential equation, discuss its restrictions, and describe what the equation represents physically. (a)
→ "r → % § & (rV ) ! 0 "t
(b)
→ →→ → " → → (rV ) % § & (rV V ) ! rg % § & sij "t
(c) For an incompressible fluid mechanics problem, the continuity equation and Cauchy’s equation provide enough equations to match the number of unknowns. (d) For an incompressible fluid mechanics problem involving a Newtonian fluid with constant properties, the continuity equation and the Navier–Stokes equation provide enough equations to match the number of unknowns. 9–105C Discuss the relationship between volumetric strain rate and the continuity equation. Base your discussion on fundamental definitions. 9–106 Repeat Example 9–17, except for the case in which the wall is moving upward at speed V. As a check, make sure that your result agrees with that of Example 9–17 when V ! 0. Nondimensionalize your velocity profile equation using the same normalization as in Example 9–17, and show that a Froude number and a Reynolds number emerge. Plot the profile w* versus x* for cases in which Fr ! 0.5 and Re ! 0.5, 1.0, and 5.0. Discuss. 9–107 For the falling oil film of Prob. 9–106, calculate the volume flow rate per unit width of oil falling down the wall # (V $L) as a function of wall speed V and the other parameters in the problem. Calculate the wall speed required such that there is no net volume flow of oil either up or down. Give your answer for V in terms of the other parameters in the problem, namely, r, m, h, and g. Calculate V for zero volume flow rate for an oil film of thickness 5.0 mm with r ! 888 kg/m3 and m ! 0.80 kg/m · s. Answer: 0.091 m/s 9–108 Consider steady, two-dimensional, incompressible flow in the xz-plane rather than in the xy-plane. Curves of constant stream function are shown in Fig. P9–108. The nonzero velocity components are (u, w). Define a stream function such that flow is from right to left in the xz-plane when c increases in the z-direction.
c = c3
z
→
→ → DV → ! # §P % rg % m § 2V (c) r Dt
9–103C List the six steps used to solve the Navier–Stokes and continuity equations for incompressible flow with constant fluid properties. (You should be able to do this without peeking at the chapter.) 9–104C For each statement, choose whether the statement is true or false and discuss your answer briefly. For each statement it is assumed that the proper boundary conditions and fluid properties are known. (a) A general incompressible flow problem with constant fluid properties has four unknowns. (b) A general compressible flow problem has five unknowns.
c = c2
x
c = c1 Streamlines
FIGURE P9–108 9–109 Consider the following steady, three-dimensional → velocity field in→Cartesian coordinates: V ! (u, v, w) ! (axz2 → → # by)i % cxyzj % (dz3 % exz2)k , where a, b, c, d, and e are constants. Under what conditions is this flow field incompressible? What are the primary dimensions of constants a, b, c, d, and e?
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468 FLUID MECHANICS
9–110 Simplify the Navier–Stokes equation as much as possible for the case of an incompressible liquid being accelerated as a rigid body in an arbitrary direction (Fig. P9–110). Gravity acts in the &z-direction. Begin with the incompressible vector form of the Navier–Stokes equation, explain how and why some terms can be simplified, and give your final result as a vector equation.
→
a
Free surface
Fluid particle
→
g
→
a
9–113 For each of the listed equations, write down the equation in vector form and decide if it is linear or nonlinear. If it is nonlinear, which term(s) make it so? (a) Incompressible continuity equation, (b) compressible continuity equation, and (c) incompressible Navier–Stokes equation. 9–114 A boundary layer is a thin region near a wall in which viscous (frictional) forces are very important due to the no-slip boundary condition. The steady, incompressible, twodimensional, boundary layer developing along a flat plate aligned with the free-stream flow is sketched in Fig. P9–114. The flow upstream of the plate is uniform, but boundary layer thickness d grows with x along the plate due to viscous effects. Sketch some streamlines, both within the boundary layer and above the boundary layer. Is d(x) a streamline? (Hint: Pay particular attention to the fact that for steady, incompressible, two-dimensional flow the volume flow rate per unit width between any two streamlines is constant.)
Liquid V∞
FIGURE P9–110
y d(x)
9–111 Simplify the Navier–Stokes equation as much as possible for the case of incompressible hydrostatics, with gravity acting in the negative z-direction. Begin with the incompressible vector form of the Navier–Stokes equation, explain how and why some terms can be simplified, and give → → your final result as a vector equation. Answer: VP " &rgk 9–112 Bob will use a computational fluid dynamics code to model steady flow of an incompressible fluid through a twodimensional sudden contraction as sketched in Fig. P9–112. Channel height changes from→H1 " 12.0 cm to H2 " 4.6 cm. → Uniform velocity V 1 " 18.5i m/s is to be specified on the left boundary of the computational domain. The CFD code uses a numerical scheme in which the stream function must be specified along all boundaries of the computational domain. As shown in Fig. P9–112, c is specified as zero along the entire bottom wall of the channel. (a) What value of c should Bob specify on the top wall of the channel? (b) How should Bob specify c on the left side of the computational domain? (c) Discuss how Bob might specify c on the right side of the computational domain.
d(x) x
Boundary layer
FIGURE P9–114 9–115E
A group of students is designing a small, round (axisymmetric), low-speed wind tunnel for their senior design project (Fig. P9–115E). Their design calls for the axial component of velocity to increase linearly in the contraction section from uz, 0 to uz, L. The air speed through the test section is to be uz, L " 120 ft/s. The length of the contraction is L " 3.0 ft, and the entrance and exit diameters of the contraction are D0 " 5.0 ft and DL " 1.5 ft, respectively. The air is at standard temperature and pressure. (a) Verify that the flow can be approximated as incompressible. (b) Generate an expression for the radial velocity component ur between z " 0 and z " L, staying in variable form. Contraction D0
y
V1 H1
H2
x
Test section
r uz, 0
uz, L z DL
c=0
FIGURE P9–112
z=0
FIGURE P9–115E
z=L
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469 CHAPTER 9
You may ignore frictional effects (boundary layers) on the walls. (c) Generate an expression for the stream function c " function of r and z. (d) Plot some streamlines and design the shape of the contraction, assuming that frictional effects along the walls of the wind tunnel contraction are negligible. 9–116 We approximate the flow of air into a vacuum cleaner’s floor attachment by the stream function c # sin 2u &V arctan in the center plane (the xy" 2pL cos 2u $ b 2(r 2 plane) in cylindrical coordinates, where L is the length of the attachment, b is the height of the attachment above the floor, # and V is the volume flow rate of air being sucked into the hose. Shown in Fig. P9–116 is a three-dimensional view with the floor in the xz-plane; we model a two-dimensional slice of the flow in the xy-plane through the centerline of the attachment. Note that we have (arbitrarily) set c " 0 along the positive x-axis (u " 0). (a) What are the primary dimensions of the given stream function? (b) Nondimensionalize the stream # function by defining c* " (2pL/V )c and r* " r/b. (c) Solve your nondimensionalized equation for r* as a function of c* and u. Use this equation to plot several nondimensional streamlines of the flow. For consistency, plot in the range &2 + x* + 2 and 0 + y* + 4, where x* " x/b and y* " y/b. (Hint: c* must be negative to yield the proper flow direction.)
•
V y L
adopt a coordinate system in which x follows the axis of the pipe. (a) Use the control volume technique of Chap. 8 to generate an expression for average velocity V as a function of the given parameters r, g, D, 'z, m, and L. (b) Use differential analysis to generate an expression for V as a function of the given parameters. Compare with your result of part (a) and discuss. (c) Use dimensional analysis to generate a dimensionless expression for V as a function of the given parameters. Construct a relationship between your -’s that matches the exact analytical expression. P1 D x ∆z
r, m
L
V
P2 a
9–119 A block slides down a long, straight, inclined wall at speed V, riding on a thin film of oil of thickness h (Fig. P9–119). The weight of the block is W, and its surface area in contact with the oil film is A. Suppose V is measured, and W, A, angle a, and viscosity m are also known. Oil film thickness h is not known. (a) Generate an exact analytical expression for h as a function of the known parameters V, A, W, a, and m. (b) Use dimensional analysis to generate a dimensionless expression for h as a function of the given parameters. Construct a relationship between your -’s that matches the exact analytical expression of part (a). →
g V
z
b
r, m x
FIGURE P9–116 9–117 Look up the definition of Poisson’s equation in one of your math textbooks or on the Internet. Write Poisson’s equation in standard form. How is Poisson’s equation similar to Laplace’s equation? How do these two equations differ? 9–118 Water flows down a long, straight, inclined pipe of diameter D and length L (Fig. P9–118). There is no forced pressure gradient between points 1 and 2; in other words, the water flows through the pipe by gravity alone, and P1 " P2 " Patm. The flow is steady, fully developed, and laminar. We
g
FIGURE P9–118
h
Floor
→
A
FIGURE P9–119
a
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CHAPTER
10
A P P R O X I M AT E S O L U T I O N S O F T H E N AV I E R – S T O K E S E Q U AT I O N
I
n this chapter we look at several approximations that eliminate term(s), reducing the Navier–Stokes equation to a simplified form that is more easily solvable. Sometimes these approximations are appropriate in a whole flow field, but in most cases, they are appropriate only in certain regions of the flow field. We first consider creeping flow, where the Reynolds number is so low that the viscous terms dominate (and eliminate) the inertial terms. Following that, we look at two approximations that are appropriate in regions of flow away from walls and wakes: inviscid flow and irrotational flow (also called potential flow). In these regions, the opposite holds; i.e., inertial terms dominate viscous terms. Finally, we discuss the boundary layer approximation, in which both inertial and viscous terms remain, but some of the viscous terms are negligible. This last approximation is appropriate at very high Reynolds numbers (the opposite of creeping flow) and near walls, the opposite of potential flow.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
■
Appreciate why approximations are necessary to solve many fluid flow problems, and know when and where such approximations are appropriate Understand the effects of the lack of inertial terms in the creeping flow approximation, including the disappearance of density from the equations Understand superposition as a method of solving potential flow problems Predict boundary layer thickness and other boundary layer properties
471
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472 FLUID MECHANICS
10–1
“Exact” solution Full Navier–Stokes equation Analysis Solution
Approximate solution Simplified Navier–Stokes equation Analysis Solution
FIGURE 10–1 “Exact” solutions begin with the full Navier–Stokes equation, while approximate solutions begin with a simplified form of the Navier–Stokes equation right from the start.
■
INTRODUCTION
In Chap. 9, we derived the differential equation of conservation of linear momentum for an incompressible Newtonian fluid with constant properties—the Navier–Stokes equation. We showed some examples of analytical solutions to this equation for simple (usually infinite) geometries, in which most of the terms in the component equations are eliminated and the resulting differential equations are analytically solvable. Unfortunately, there aren’t very many known analytical solutions available in the literature; in fact, we can count the number of such solutions on the fingers of a few students. The vast majority of practical fluid mechanics problems cannot be solved analytically and require either (1) further approximations or (2) computer assistance. We consider option 1 here; option 2 is discussed in Chap. 15. For simplicity, we consider only incompressible flow of Newtonian fluids in this chapter. We emphasize first that the Navier–Stokes equation itself is not exact, but rather is a model of fluid flow that involves several inherent approximations (Newtonian fluid, constant thermodynamic and transport properties, etc.). Nevertheless, it is an excellent model and is the foundation of modern fluid mechanics. In this chapter we distinguish between “exact” solutions and approximate solutions (Fig. 10–1). The term exact is used when the solution starts with the full Navier–Stokes equation. The solutions discussed in Chap. 9 are exact solutions because we begin each of them with the full form of the equation. Some terms are eliminated in a specific problem due to the specified geometry or other simplifying assumptions in the problem. In a different solution, the terms that get eliminated may not be the same ones, but depend on the geometry and assumptions of that particular problem. We define an approximate solution, on the other hand, as one in which the Navier–Stokes equation is simplified in some region of the flow before we even start the solution. In other words, term(s) are eliminated a priori depending on the class of problem, which may differ from one region of the flow to another. For example, we have already discussed one approximation, namely, fluid statics (Chap. 3). This can be considered to be an approximation of the Navier–Stokes equation in a region of the flow field where the fluid velocity is not necessarily zero, but the fluid is nearly stagnant, and we neglect all terms involving velocity. In this approximation, the Navier–Stokes equation → → reduces to just two terms, pressure and gravity, i.e., §P ! rg . The approximation is that the inertial and viscous terms in the Navier–Stokes equation are negligibly small compared to the pressure and gravity terms. Although approximations render the problem more tractable, there is a danger associated with any approximate solution. Namely, if the approximation is not appropriate to begin with, the solution will be incorrect—even if we perform all the mathematics correctly. Why? Because we start with equations that do not apply to the problem at hand. For example, we may solve a problem using the creeping flow approximation and obtain a solution that satisfies all assumptions and boundary conditions. However, if the Reynolds number of the flow is too high, the creeping flow approximation is inappropriate from the start, and our solution (regardless of how proud of it we may be) is not physically correct. Another common mistake is to
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473 CHAPTER 10 Supply tank
Receiving tank
Fluid statics region
Boundary layer region
Irrotational flow region
FIGURE 10–2 A particular approximation of the Navier–Stokes equation is appropriate only in certain regions of the flow field; other approximations may be appropriate in other regions of the flow field.
Fluid statics region
Full Navier– Stokes region
assume irrotational flow in regions of the flow where the assumption of irrotationality is not appropriate. The bottom line is that we must be very careful of the approximations we apply, and we should always verify and justify our approximations wherever possible. Finally, we stress that in most practical fluid flow problems, a particular approximation may be appropriate in a certain region of the flow field, but not in other regions, where a different approximation may perhaps be more appropriate. Figure 10–2 illustrates this point qualitatively for flow of a liquid from one tank to another. The fluid statics approximation is appropriate in a region of the supply tank far away from the connecting pipe, and to a lesser extent in the receiving tank. The irrotational flow approximation is appropriate near the inlet to the connecting pipe and through the middle portion of the pipe where strong viscous effects are absent. Near the walls, the boundary layer approximation is appropriate. The flow in some regions does not meet the criteria for any approximations, and the full Navier– Stokes equation must be solved there (e.g., near the pipe outlet in the receiving tank). How do we determine if an approximation is appropriate? We do this by comparing the orders of magnitude of the various terms in the equations of motion to see if any terms are negligibly small compared to other terms.
10–2
■
NONDIMENSIONALIZED EQUATIONS OF MOTION
Our goal in this section is to nondimensionalize the equations of motion so that we can properly compare the orders of magnitude of the various terms in the equations. We begin with the incompressible continuity equation, →
→
§ $V !0
(10–1)
and the vector form of the Navier–Stokes equation, valid for incompressible flow of a Newtonian fluid with constant properties, →
→
→ → → → → DV "V → r ! rc # aV $ §bV d ! %§P # rg # m§ 2V Dt "t
(10–2)
We introduce in Table 10–1 some characteristic (reference) scaling parameters that are used to nondimensionalize the equations of motion.
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474 FLUID MECHANICS
TA B L E 1 0 – 1 Scaling parameters used to nondimensionalize the continuity and momentum equations, along with their primary dimensions
∂ ∂ ∂ , , b ∂x ∂y ∂z
=a
∂
∂
,
∂
L
→
*
a
→
g g ! g
b
1 ∂ 1 ∂ 1 ∂ = , , b= a L ∂r* r* ∂u ∂z* L
→
→*
*
*
(10–3)
→
§ ! L§
→
rV 2 → * → * → * §* → aV $ § bV r(V $ § )V ! r aVV $ b VV * ! L L →
∂ 1 ∂ , , ∂u r r z L∂ a b L a b L∂ a b L L L
→
V V ! V →
Notice that we define the nondimensional pressure variable in terms of a pressure difference, based on our discussion about pressure versus pressure differences in Chap. 9. Each of the starred quantities in Eq. 10–3 is nondimensional. For example, although each component of the gradient operator → → § has dimensions of {L%1}, each component of §* has dimensions of {1} (Fig. 10–3). We substitute Eq. 10–3 into Eqs. 10–1 and 10–2, treating each → → → → term carefully. For example, § ! §*/L and V ! VV *, so the advective acceleration term in Eq. 10–2 becomes
L
∂ 1 ∂ ∂ , , , b ∂r r ∂u ∂z ∂
x x ! L
→*
P % P& P ! P0 % P&
→ →
→
*
We perform similar algebra on each term in Eqs. 10–1 and 10–2. Equation 10–1 is rewritten in terms of nondimensional variables as →
∆
∆
=
{L} {Lt%1} {t%1} {mL%1t%2} {Lt%2}
*
1 ∂ 1 ∂ 1 , a b= L ∂x* ∂y* ∂z* L
= a
Characteristic length Characteristic speed Characteristic frequency Reference pressure difference Gravitational acceleration
t ! ft
Cylindrical coordinates →
L V f P 0 % P& g
→
aL∂ a xb L∂ a y b L∂ az b b L
=
,
Primary Dimensions
*
∆
∆
=
Description
We then define several nondimensional variables and one nondimensional operator based on the scaling parameters in Table 10–1,
Cartesian coordinates →
Scaling Parameter
FIGURE 10–3 The gradient operator is nondimensionalized by Eq. 10–3, regardless of our choice of coordinate system.
V →* → * § $V !0 L
*
After dividing both sides by V/L to make the equation dimensionless, we get Nondimensionalized continuity:
→
→
§* $ V * ! 0
(10–4)
Similarly, Eq. 10–2 is rewritten as →
rVf
→ P0 % P& → * * mV "V * rV 2 → * → * → * → # aV $ § bV ! % § P # rgg * # 2 §*2V * * L L "t L
which, after multiplication by the collection of constants L/(rV 2) to make all the terms dimensionless, becomes →
→ → → → P0 % P& → * * m fL "V * gL → c d * # aV * $ §*bV * ! % c d § P # c 2dg * # c d§*2V * 2 V "t rVL rV V
(10–5)
Each of the terms in square brackets in Eq. 10–5 is a nondimensional grouping of parameters—a Pi group (Chap. 7). With the help of Table 7–5, we name each of these dimensionless parameters: The one on the left is the Strouhal number, St ! fL/V; the first one on the right is the Euler number,
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475 CHAPTER 10
Eu ! (P0 % P&)/rV 2; the second one on the right is the reciprocal of the square of the Froude number, Fr2 ! V 2/gL; and the last one is the reciprocal of the Reynolds number, Re ! rVL/m. Equation 10–5 thus becomes Nondimensionalized Navier–Stokes: →
→ → → → → "V * 1 → 1 [St] * # (V * $ §*)V * ! %[Eu]§*P * # c 2dg * # c d§*2V * Re "t Fr
(10–6)
Before we discuss specific approximations in detail, there is much to comment about the nondimensionalized equation set consisting of Eqs. 10–4 and 10–6: • The nondimensionalized continuity equation contains no additional dimensionless parameters. Hence, Eq. 10–4 must be satisfied as is—we cannot simplify continuity further, because all the terms are of the same order of magnitude. • The order of magnitude of the nondimensional variables is unity if they are nondimensionalized using a length, speed, frequency, etc., that are → → characteristic of the flow field. Thus, t* ! 1, x * ! 1, V * ! 1, etc., where we use the→notation ! to denote order of magnitude. It follows that terms → → → like (V * $ §*)V * and §*P * in Eq. 10–6 are also order of magnitude unity and are the same order of magnitude as each other. Thus, the relative importance of the terms in Eq. 10–6 depends only on the relative magnitudes of the dimensionless parameters St, Eu, Fr, and Re. For example, if St and Eu are of order 1, but Fr and Re are very large, we may consider ignoring the gravitational and viscous terms in the Navier–Stokes equation. • Since there are four dimensionless parameters in Eq. 10–6, dynamic similarity between a model and a prototype requires all four of these to be the same for the model and the prototype (Stmodel ! Stprototype, Eumodel ! Euprototype, Frmodel ! Frprototype, and Remodel ! Reprototype), as illustrated in Fig. 10–4. • If the flow is steady, then f ! 0 and the Strouhal number drops out of the list of dimensionless parameters (St ! 0). The first term on the left side of → Eq. 10–6 then disappears, as does its corresponding unsteady term "V /"t in Eq. 10–2. If the characteristic frequency f is very small such that St '' 1, the flow is called quasi-steady. This means that at any moment of time (or at any phase of a slow periodic cycle), we can solve the problem as if the flow were steady, and the unsteady term in Eq. 10–6 again drops out. • The effect of gravity is important only in flows with free-surface effects (e.g., waves, ship motion, spillways from hydroelectric dams, flow of rivers). For many engineering problems there is no free surface (pipe flow, fully submerged flow around a submarine or torpedo, automobile motion, flight of airplanes, birds, insects, etc.). In such cases, the only effect of gravity on the flow dynamics is a hydrostatic pressure distribution in the vertical direction superposed on the pressure field due to the fluid flow. In other words, For flows without free-surface effects, gravity does not affect the dynamics of the flow—its only effect is to superpose a hydrostatic pressure on the dynamic pressure field.
Prototype Stprototype, Euprototype, Frprototype, Reprototype →
gp
P∞, p Vp
fp
P0, p
Lp
Model Stmodel, Eumodel, Frmodel, Remodel
→
fm
Vm
P∞, m P0, m
gm Lm
FIGURE 10–4 For complete dynamic similarity between prototype (subscript p) and model (subscript m), the model must be geometrically similar to the prototype, and (in general) all four dimensionless parameters, St, Eu, Fr, and Re, must match.
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476 FLUID MECHANICS
• We define a modified pressure P( that absorbs the effect of hydrostatic pressure. For the case in which z is defined vertically upward (opposite to the direction of the gravity vector), and in which we define some arbitrary reference datum plane at z ! 0,
V P' P
→
g
z x
V
→
z
→
x
P
→
(10–7) →
(b)
FIGURE 10–5 Pressure and modified pressure distribution on the right face of a fluid element in Couette flow between two infinite, parallel, horizontal plates: (a) z ! 0 at the bottom plate, and (b) z ! 0 at the top plate. The modified pressure P( is constant, but the actual pressure P is not constant in either case. The shaded area in (b) represents the hydrostatic pressure component.
FIGURE 10–6 The slow flow of a very viscous liquid like honey is classified as creeping flow.
(10–8)
With P replaced by P(, and with the gravity term removed from Eq. 10–2, the Froude number drops out of the list of dimensionless parameters. The advantage is that we can solve a form of the Navier–Stokes equation that has no gravity term. After solving the Navier–Stokes equation in terms of modified pressure P(, it is a simple matter to add back the hydrostatic pressure distribution using Eq. 10–7. An example is shown in Fig. 10–5 for the case of two-dimensional Couette flow. Modified pressure is often used in computational fluid dynamics (CFD) codes to separate gravitational effects (hydrostatic pressure in the vertical direction) from fluid flow (dynamic) effects. Note that modified pressure should not be used in flows with free-surface effects. Now we are ready to make some approximations, in which we eliminate one or more of the terms in Eq. 10–2 by comparing the relative magnitudes of the dimensionless parameters associated with the corresponding terms in Eq. 10–6.
10–3
Direct to you from the lovely Stokes Valley
→
→ → → → → DV "V ! r c # (V $ §)V d ! %§ P( # m§ 2V r Dt "t
P'
g
P( ! P # rgz
The idea is to replace the two terms %§P # rg in Eq. 10–2 with one term → %§P( using the modified pressure of Eq. 10–7. The Navier–Stokes equation (Eq. 10–2) is written in modified form as
(a) Hydrostatic pressure
Modified pressure:
■
THE CREEPING FLOW APPROXIMATION
Our first approximation is the class of fluid flow called creeping flow. Other names for this class of flow include Stokes flow and low Reynolds number flow. As the latter name implies, these are flows in which the Reynolds number is very small (Re '' 1). By inspection of the definition of the Reynolds number, Re ! rVL/m, we see that creeping flow is encountered when either r, V, or L is very small or viscosity is very large (or some combination of these). You encounter creeping flow when you pour syrup (a very viscous liquid) on your pancakes or when you dip a spoon into a jar of honey (also very viscous) to add to your tea (Fig. 10–6). Another example of creeping flow is all around us and inside us, although we can’t see it, namely, flow around microscopic organisms. Microorganisms live their entire lives in the creeping flow regime since they are very small, their size being of order one micron (1 )m ! 10%6 m), and they move very slowly, even though they may move in air or swim in water with a viscosity that can hardly be classified as “large” (mair ≅ 1.8 * 10%5 N · s/m2 and mwater ≅ 1.0 * 10%3 N · s/m2 at room temperature). Figure 10–7 shows a Salmonella bacterium swimming through water. The bacterium’s body is only about 1 )m long; its flagella (hairlike tails) extend several microns behind the body and serve as its propulsion mechanism. The Reynolds number associated with its motion is much smaller than 1.
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477 CHAPTER 10
Creeping flow also occurs in the flow of lubricating oil in the very small gaps and channels of a lubricated bearing. In this case, the speeds may not be small, but the gap size is very small (on the order of tens of microns), and the viscosity is relatively large (moil ! 1 N · s/m2 at room temperature). For simplicity, we assume that gravitational effects are negligible, or that they contribute only to a hydrostatic pressure component, as discussed previously. We also assume either steady flow or oscillating flow, with a Strouhal number of order unity (St ! 1) or smaller, so that the unsteady → */"t* is orders of magnitude smaller than the visacceleration term→[St] "V → number is very small). The advective cous term [1/Re]§*2V * (the Reynolds → → → term in Eq. 10–6 is of order 1, (V * $ §*)V * ! 1, so this term drops out as well. Thus, we ignore the entire left side of Eq. 10–6, which reduces to → → 1 [Eu]§ *P * " c d§ * 2V * Re
Creeping flow approximation:
(10–9)
In words, pressure forces in the flow (left side) must be large enough to balance the (relatively) large viscous forces on the right side. However, since the nondimensional variables in Eq. 10–9 are of order 1, the only way for the two sides to balance is if Eu is of the same order of magnitude as 1/Re. Equating these, [Eu] !
After some algebra, Pressure scale for creeping flow:
P0 % P& rV
2
m 1 !c d! Re rVL P0 % P& !
mV L
(10–10)
Equation 10–10 reveals two interesting properties of creeping flow. First, we are used to inertially dominated flows, in which pressure differences scale like rV 2 (e.g., the Bernoulli equation). Here, however, pressure differences scale like mV/L instead, since creeping flow is a viscously dominated flow. In fact, all the inertial terms of the Navier–Stokes equation disappear in creeping flow. Second, density has completely dropped out as a parameter in the Navier–Stokes equation (Fig. 10–8). We see this more clearly by writing the dimensional form of Eq. 10–9, Approximate Navier–Stokes equation for creeping flow:
→
→
§ P " m§ 2V
FIGURE 10–7 The bacterium Salmonella abortusequi swimming through water. From Comparative Physiology Functional Aspects of Structural Materials: Proceedings of the International Conference on Comparative Physiology, Ascona, 1974, published by North-Holland Pub. Co., 1975.
→
(10–11)
Alert readers may point out that density still has a minor role in creeping flow. Namely, it is needed in the calculation of the Reynolds number. However, once we have determined that Re is very small, density is no longer needed since it does not appear in Eq. 10–11. Density also pops up in the hydrostatic pressure term, but this effect is usually negligible in creeping flow, since the vertical distances involved are often measured in millimeters or micrometers. Besides, if there are no free-surface effects, we can use modified pressure instead of physical pressure in Eq. 10–11. Let’s discuss the lack of inertia terms in Eq. 10–11 in somewhat more detail. You rely on inertia when you swim (Fig. 10–9). For example, you take a stroke, and then you are able to glide for some distance before you need to take another stroke. When you swim, the inertial terms in the Navier–Stokes equation are much larger than the viscous terms, since the
+P "
→
m+ 2V
Density? What is density?
FIGURE 10–8 In the creeping flow approximation, density does not appear in the momentum equation.
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478 FLUID MECHANICS
FIGURE 10–9 A person swims at a very high Reynolds number, and inertial terms are large; thus the person is able to glide long distances without moving.
Reynolds number is very large. (Believe it or not, even extremely slow swimmers move at very large Reynolds numbers!) For microorganisms swimming in the creeping flow regime, however, there is negligible inertia, and thus no gliding is possible. In fact, the lack of inertial terms in Eq. 10–11 has a substantial impact on how microorganisms are designed to swim. A flapping tail like that of a dolphin would get them nowhere. Instead, their long, narrow tails (flagella) undulate in a sinusoidal motion to propel them forward, as illustrated in Fig. 10–10 for the case of a sperm. Without any inertia, the sperm does not move unless his tail is moving. The instant his tail stops, the sperm stops moving. If you have ever seen a video clip of swimming sperm or other microorganisms, you may have noticed how hard they have to work just to move a short distance. That is the nature of creeping flow, and it is due to the lack of inertia. Careful study of Fig. 10–10 reveals that the sperm’s tail has completed approximately two complete undulation cycles, yet the sperm’s head has moved to the left by only about two head lengths. It is very difficult for us humans to imagine moving in creeping flow conditions, since we are so used to the effects of inertia. Some authors have suggested that you imagine trying to swim in a vat of honey. We suggest instead that you go to a fast-food restaurant where they have a children’s play area and watch a child play in a pool of plastic spheres (Fig. 10–11). When the child tries to “swim” among the balls (without touching the walls or the bottom), he or she can move forward only by certain snakelike wriggling body motions. The instant the child stops wriggling, all motion stops, since there is negligible inertia. The child must work very hard to move forward a short distance. There is a weak analogy between a child “swimming” in this kind of situation and a microorganism swimming in creeping flow conditions. Now let’s discuss the lack of density in Eq. 10–11. At high Reynolds numbers, the aerodynamic drag on an object increases proportionally with r. (Denser fluids exert more pressure force on the body as the fluid impacts the body.) However, this is actually an inertial effect, and inertia is negligible in creeping flow. In fact, aerodynamic drag cannot even be a function of density, since density has disappeared from the Navier–Stokes equation. Example 10–1 illustrates this situation through the use of dimensional analysis.
EXAMPLE 10–1
10 mm
FIGURE 10–10 A sperm of the sea squirt Ciona swimming in seawater; flash photographs at 200 frames per second. Courtesy Charlotte Omoto and Charles J. Brokaw. Used by permission.
Drag on an Object in Creeping Flow
Since density has vanished from the Navier–Stokes equation, aerodynamic drag on an object in creeping flow is a function only of its speed V, some characteristic length scale L of the object, and fluid viscosity m (Fig. 10–12). Use dimensional analysis to generate a relationship for FD as a function of these independent variables.
SOLUTION We are to use dimensional analysis to generate a functional relationship between FD and variables V, L, and m. Assumptions 1 We assume Re '' 1 so that the creeping flow approximation applies. 2 Gravitational effects are irrelevant. 3 No parameters other than those listed in the problem statement are relevant to the problem.
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479 CHAPTER 10
Analysis We follow the step-by-step method of repeating variables discussed in Chap. 7; the details are left as an exercise. There are four parameters in this problem (n ! 4). There are three primary dimensions: mass, length, and time, so we set j ! 3 and use independent variables V, L, and m as our repeating variables. We expect only one Pi since k ! n % j ! 4 % 3 ! 1, and that Pi must equal a constant. The result is
FD ! constant " MVL Thus, we have shown that for creeping flow around any three-dimensional object, the aerodynamic drag force is simply a constant multiplied by mVL. Discussion This result is significant, because all that is left to do is find the constant, which is a function only of the shape of the object.
Drag on a Sphere in Creeping Flow
As shown in Example 10–1, the drag force FD on a three-dimensional object of characteristic dimension L moving under creeping flow conditions at speed V through a fluid with viscosity m is FD ! constant $ mVL. Dimensional analysis cannot predict the value of the constant, since it depends on the shape and orientation of the body in the flow field. For the particular case of a sphere, Eq. 10–11 can be solved analytically. The details are beyond the scope of this text, but can be found in graduatelevel fluid mechanics books (White, 1991; Panton, 1996). It turns out that the constant in the drag equation is equal to 3p if L is taken as the sphere’s diameter D (Fig. 10–13). Drag force on a sphere in creeping flow:
FD ! 3pmVD
m V
FD
L
(10–12)
As a side note, two-thirds of this drag is due to viscous forces and the other one-third is due to pressure forces. This confirms that the viscous terms and the pressure terms in Eq. 10–11 are of the same order of magnitude, as mentioned previously. EXAMPLE 10–2
FIGURE 10–11 A child trying to move in a pool of plastic balls is analogous to a microorganism trying to propel itself without the benefit of inertia.
Terminal Velocity of a Particle from a Volcano
A volcano has erupted, spewing stones, steam, and ash several thousand feet into the atmosphere (Fig. 10–14). After some time, the particles begin to settle to the ground. Consider a nearly spherical ash particle of diameter 50 mm, falling in air whose temperature is %50°C and whose pressure is 55 kPa. The density of the particle is 1240 kg/m3. Estimate the terminal velocity of this particle at this altitude.
SOLUTION We are to estimate the terminal velocity of a falling ash particle. Assumptions 1 The Reynolds number is very small (we will need to verify this assumption after we obtain the solution). 2 The particle is spherical. Properties At the given temperature and pressure, the ideal gas law gives r ! 0.8588 kg/m3. Since viscosity is a very weak function of pressure, we use the value at %50°C and atmospheric pressure, m ! 1.474 * 10%5 kg/m · s. Analysis We treat the problem as quasi-steady. Once the falling particle has reached its terminal velocity, the net downward force (weight) balances
FIGURE 10–12 For creeping flow over a threedimensional object, the aerodynamic drag on the object does not depend on density, but only on speed V, some characteristic size of the object L, and fluid viscosity m.
m V
D FD
FIGURE 10–13 The aerodynamic drag on a sphere of diameter D in creeping flow is equal to 3pmVD.
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480 FLUID MECHANICS
the net upward force (aerodynamic drag # buoyancy), as illustrated in Fig. 10–15. V Terminal velocity
Fdown ! W ! p
Downward force:
D3 r g 6 particle
(1)
The aerodynamic drag force acting on the particle is obtained from Eq. 10–12, and the buoyancy force is the weight of the displaced air. Thus,
Upward force:
Fup ! FD # Fbuoyancy ! 3pmVD # p
D3 r g 6 air
(2)
We equate Eqs. 1 and 2, and solve for terminal velocity V,
FIGURE 10–14 Small ash particles spewed from a volcanic eruption settle slowly to the ground; the creeping flow approximation is reasonable for this type of flow field.
!
D2 (r % r air)g 18m particle (50 * 10 %6 m)2 [(1240 % 0.8588) kg,m3](9.81 m,s2) 18(1.474 * 10 %5 kg,m $ s)
! 0.115 m,s Finally, we verify that the Reynolds number is small enough that creeping flow is an appropriate approximation,
FD
rair, mair
Re !
D rparticle
V!
Fbuoyancy V
W
FIGURE 10–15 A particle falling at a steady terminal velocity has no acceleration; therefore, its weight is balanced by aerodynamic drag and the buoyancy force acting on the particle.
r airVD (0.8588 kg,m3)(0.115 m,s)(50 * 10 %6 m) ! ! 0.335 m 1.474 * 10 %5 kg,m $ s
Thus the Reynolds number is less than 1, but certainly not much less than 1. Discussion Although the equation for creeping flow drag on a sphere (Eq. 10–12) was derived for a case with Re '' 1, it turns out that the approximation is reasonable up to Re ≅ 1. A more involved calculation, including a Reynolds number correction and a correction based on the mean free path of air molecules, yields a terminal velocity of 0.110 m/s (Heinsohn and Cimbala, 2003); the error of the creeping flow approximation is less than 5 percent.
A consequence of the disappearance of density from the equations of motion for creeping flow is clearly seen in Example 10–2. Namely, air density is not important in any calculations except to verify that the Reynolds number is small. (Note that since rair is so small compared to rparticle, the buoyancy force could have been ignored with negligible loss of accuracy.) Suppose instead that the air density were one-half of the actual density in Example 10–2, but all other properties were unchanged. The terminal velocity would be the same (to three significant digits), except that the Reynolds number would be smaller by a factor of 2. Thus, The terminal velocity of a dense, small particle in creeping flow conditions is independent of fluid density, but highly dependent on fluid viscosity.
Since the viscosity of air varies with altitude by only about 25 percent, a small particle settles at nearly constant speed regardless of elevation, even though the air density increases by more than a factor of 10 as the particle falls from an altitude of 50,000 ft (15,000 m) to sea level.
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481 CHAPTER 10
For nonspherical three-dimensional objects, the creeping flow aerodynamic drag is still given by FD ! constant $ mVL; however, the constant is not 3p, but depends on both the shape and orientation of the body. The constant can be thought of as a kind of drag coefficient for creeping flow.
10–4
■
APPROXIMATION FOR INVISCID REGIONS OF FLOW
There is much confusion in the fluid mechanics literature about the word inviscid and the phrase inviscid flow. The apparent meaning of inviscid is not viscous. Inviscid flow would then seem to refer to flow of a fluid with no viscosity. However, that is not what is meant by the phrase “inviscid flow”! All fluids of engineering relevance have viscosity, regardless of the flow field. Authors who use the phrase inviscid flow actually mean flow of a viscous fluid in a region of the flow in which net viscous forces are negligible compared to pressure and/or inertial forces (Fig. 10–16). Some authors use the phrase “frictionless flow” as a synonym of inviscid flow. This causes more confusion, because even in regions of the flow where net viscous forces are negligible, friction still acts on fluid elements, and there may still be significant viscous stresses. It’s just that these stresses cancel each other out, leaving no significant net viscous force on fluid elements. It can be shown that significant viscous dissipation may also be present in such regions. As is discussed in Section 10–5, fluid elements in an irrotational region of the flow also have negligible net viscous forces—not because there is no friction, but because the frictional (viscous) stresses cancel each other out. Because of the confusion caused by the terminology, the present authors discourage use of the phrases “inviscid flow” and “frictionless flow.” Instead, we advocate use of the phrases inviscid regions of flow or regions of flow with negligible net viscous forces. Regardless of the terminology used, if net viscous forces are very small compared to inertial and/or pressure forces, the last term on the right side of Eq. 10–6 is negligible. This is true only if 1/Re is small. Thus, inviscid regions of flow are regions of high Reynolds number—the opposite of creeping flow regions. In such regions, the Navier–Stokes equation (Eq. 10–2) loses its viscous term and reduces to the Euler equation, →
Euler equation:
→ → → → "V → rc # (V $ § )V d ! %§ P # rg "t
(10–13)
The Euler equation is simply the Navier–Stokes equation with the viscous term neglected; it is an approximation of the Navier–Stokes equation. Because of the no-slip condition at solid walls, frictional forces are not negligible in a region of flow very near a solid wall. In such a region, called a boundary layer, the velocity gradients normal to the wall are large enough to offset the small value of 1/Re. An alternate explanation is that the characteristic length scale of the body (L) is no longer the most appropriate length scale inside a boundary layer and must be replaced by a much smaller length scale associated with the distance from the wall. When we define the Reynolds number with this smaller length scale, Re is no longer large, and the viscous term in the Navier–Stokes equation cannot be neglected.
r, m
Streamlines →
r C ∂V + (V • ∇)V D = –∇P + rg + m∇2V ∂t →
→ →
→
→
→
negligible
FIGURE 10–16 An inviscid region of flow is a region where net viscous forces are negligible compared to inertial and/or pressure forces because the Reynolds number is large; the fluid itself is still a viscous fluid.
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482 FLUID MECHANICS Euler equation valid
A similar argument can be made in the wake of a body, where velocity gradients are relatively large and the viscous terms are not negligible compared to inertial terms (Fig. 10–17). In practice, therefore, it turns out that The Euler equation approximation is appropriate in high Reynolds number regions of the flow, where net viscous forces are negligible, far away from walls and wakes.
Euler equation not valid
FIGURE 10–17 The Euler equation is an approximation of the Navier–Stokes equation, appropriate only in regions of the flow where the Reynolds number is large and where net viscous forces are negligible compared to inertial and/or pressure forces.
The term that is neglected in the Euler approximation of the Navier– → Stokes equation (m§2V ) is the term that contains the highest-order derivatives of velocity. Mathematically, loss of this term reduces the number of boundary conditions that we can specify. It turns out that when we use the Euler equation approximation, we cannot specify the no-slip boundary condition at solid walls, although we still specify that fluid cannot flow through the wall (the wall is impermeable). Solutions of the Euler equation are therefore not physically meaningful near solid walls, since flow is allowed to slip there. Nevertheless, as we show in Section 10–6, the Euler equation is often used as the first step in a boundary layer approximation. Namely, the Euler equation is applied over the whole flow field, including regions close to walls and wakes, where we know the approximation is not appropriate. Then, a thin boundary layer is inserted in these regions as a correction to account for viscous effects.
Derivation of the Bernoulli Equation in Inviscid Regions of Flow
In Chap. 5, we derived the Bernoulli equation along a streamline. Here we show an alternative derivation based on the Euler equation. For simplicity, we assume steady incompressible flow. The advective term in Eq. 10–13 can be rewritten through use of a vector identity, →
→ →
→
→ → → V2 b % V * (§ * V ) 2
(V $ §)V ! § a
Vector identity:
(10–14)
→
where V is the magnitude of vector V . We recognize the second term in → parentheses on the right side as the vorticity vector z (see Chap. 4); thus, z = vertical distance
→
→
k = unit vector in z-direction →
(z) =
∂z → ∂z → ∂z → → i + j+ k=k ∂x ∂y ∂z
∆
0
0
1
→
g
→
∆
→
→
∆
→
Thus, g = –gk = –g z = (–gz)
→
→ → V2 b %V * z 2
and an alternate form of the steady Euler equation is written as →
→ → V2 §P → → P → # g ! § a% b # g §a b % V * z ! % r r 2 →
(10–15)
where we have divided each term by the density and moved r under the gradient operator, since density is constant in an incompressible flow. We make the further assumption that gravity acts only in the %z-direction (Fig. 10–18), so that →
FIGURE 10–18 When gravity acts in the %z-direction, gravity vector g→ can be written → as §(%gz).
→ →
(V $ §)V ! §a
→
→
→
g ! %gk ! %g§ z ! § (%gz)
(10–16) →
where we have used the fact that the gradient of coordinate z is unit vector k in the z-direction. Note also that g is a constant, which allows us to move it (and the negative sign) within the gradient operator. We substitute Eq.
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483 CHAPTER 10
10–16 into Eq. 10–15, and rearrange by combining three terms within one gradient operator,
Streamline (10–17) →
→
→
→
From→the definition of the cross product of two vectors, C ! A * B , the → → vector C is perpendicular to both A and B . The left side of Eq. 10–17 must therefore be a vector everywhere perpendicular to the local velocity vector → → V , since V appears in the cross product on the right side of Eq. 10–17. Now consider flow along a three-dimensional streamline (Fig. 10–19), which by definition is everywhere→ parallel to the local velocity vector. At every point along the streamline, §(P/r # V 2/2 # gz) must be perpendicular to the streamline. Now dust off your vector algebra book and recall that the gradient of a scalar points in the direction of maximum increase of the scalar. Furthermore, the gradient of a scalar is a vector that points perpendicular to an imaginary surface on which the scalar is constant. Thus, we argue that the scalar (P/r # V 2/2 # gz) must→be constant along a streamline. This is true even if the flow is rotational (z - 0). Thus, we have derived a version of the steady incompressible Bernoulli equation, appropriate in regions of flow with negligible net viscous forces, i.e., in so-called inviscid regions of flow. Steady incompressible Bernoulli equation in inviscid regions of flow: P V # # gz ! C ! constant along streamlines r 2
→
z, k
a
P r
+
V 2
2
+ gzb
→
y, j →
x, i
FIGURE 10–19 → Along a streamline, §(P/r # V 2/2 # gz) is a vector everywhere perpendicular to the streamline; hence, P/r # V 2/2 # gz is constant along the streamline. P V2 + + gz = C r 2 C = C1
uu uu = vr
C = C2
2
(10–18)
Note that the Bernoulli “constant” C in Eq. 10–18 is constant only along a streamline; the constant may change from streamline to streamline. You may be wondering if it is physically possible to have a rotational region of flow that is also inviscid, since rotationality is usually caused by viscosity. Yes, it is possible, and we give one simple example—solid body rotation (Fig. 10–20). Although the rotation may have been generated by viscous forces, a region of flow in solid body rotation has no shear and no net viscous force; it is an inviscid region of flow, even though it is also rotational. As a consequence of the rotational nature of this flow field, Eq. 10–18 applies to every streamline in the flow, but the Bernoulli constant C differs from streamline to streamline, as illustrated in Fig. 10–20.
EXAMPLE 10–3
→
V
∆
→ P → → V2 # gzb ! V * z §a # r 2
→
z
Pressure Field in Solid Body Rotation
A fluid is rotating as a rigid body (solid body rotation) around the z-axis as illustrated in Fig. 10–20. The steady incompressible velocity field is given by ur ! 0, uu ! vr, and uz ! 0. The pressure at the origin is equal to P0. Calculate the pressure field everywhere in the flow, and determine the Bernoulli constant along each streamline.
SOLUTION For a given velocity field, we are to calculate the pressure field and the Bernoulli constant along each streamline.
C = C3 r
FIGURE 10–20 Solid body rotation is an example of an inviscid region of flow that is also rotational. The Bernoulli constant C differs from streamline to streamline but is constant along any particular streamline.
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484 FLUID MECHANICS
Assumptions 1 The flow is steady and incompressible. 2 Since there is no flow in the z- (vertical) direction, a hydrostatic pressure distribution exists in the vertical direction. 3 The entire flow field is approximated as an inviscid region of flow since viscous forces are zero. 4 There is no variation of any flow variable in the u-direction. Analysis Equation 10–18 can be applied directly because of assumption 3,
1 P ! rC % rV 2 % rgz 2
Bernoulli equation:
(1)
where C is the Bernoulli constant that varies with radius as illustrated in Fig. 10–20. At any radial location r, V 2 ! v2r 2, and Eq. 1 becomes
P ! rC % r
v 2r 2 % rgz 2
(2)
At the origin (r ! 0, z ! 0), the pressure is equal to P0 (from the given boundary condition). Thus we calculate C ! C0 at the origin (r ! 0),
P0 ! rC 0
Boundary condition at the origin:
→
C0 !
P0 r
But how can we find C at an arbitrary radial location r? Equation 2 alone is insufficient since both C and P are unknowns. The answer is that we must use the Euler equation. Since there is no free surface, we employ the modified pressure of Eq. 10–7. The r-component of the Euler equation in cylindrical coordinates reduces to
u 2u "P( ! r ! rv 2r r "r
r-component of Euler equation:
where we have substituted the given value of uu. Since hydrostatic pressure is already included in the modified pressure, P( is not a function of z. By assumptions 1 and 4, respectively, P( is also not a function of t or u. Thus P( is a function of r only, and we replace the partial derivative in Eq. 3 with a total derivative. Integration yields
5 4.5
v 2r 2 # B1 2
(4)
where B1 is a constant of integration. At the origin, modified pressure P( is equal to actual pressure P, since z ! 0 there. Thus, constant B1 is found by applying the known pressure boundary condition at the origin. It turns out that B1 is equal to P0. We now convert Eq. 4 back to actual pressure using Eq. 10–7, P ! P( % rgz,
3.5
rv2R2
P( ! r
Modified pressure field:
4
P – P0
(3)
3 2.5 2 1.5
Actual pressure field:
1 0.5 0 0
0.5
1
1.5 2 r/R
2.5
FIGURE 10–21 Nondimensional pressure as a function of nondimensional radial location at zero elevation for a fluid in solid body rotation.
3
P!R
V2r2 # P0 $ Rgz 2
(5)
At the reference datum plane (z ! 0), we plot nondimensional pressure as a function of nondimensional radius, where some arbitrary radial location r ! R is chosen as a characteristic length scale in the flow (Fig. 10–21). The pressure distribution is parabolic with respect to r. Finally, we equate Eqs. 2 and 5 to solve for C,
Bernoulli constant as a function of r:
C!
P0 # V2r2 R
(6)
At the origin, C ! C0 ! P0/r, which agrees with our previous calculation.
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485 CHAPTER 10
Discussion For a fluid in solid body rotation, the Bernoulli constant increases as r 2. This is not surprising, since fluid particles move faster at larger values of r, and thus they possess more energy. In fact, Eq. 5 reveals that pressure itself increases as r 2. Physically, the pressure gradient in the radial direction provides the centripetal force necessary to keep fluid particles revolving about the origin.
10–5
■
THE IRROTATIONAL FLOW APPROXIMATION
As was pointed out in Chap. 4, there are regions of flow in which fluid particles have no net rotation; these regions are called irrotational. You must keep in mind that the assumption of irrotationality is an approximation, which may be appropriate in some regions of a flow field, but not in other regions (Fig. 10–22). In general, inviscid regions of flow far away from solid walls and wakes of bodies are also irrotational, although as pointed out previously, there are situations in which an inviscid region of flow may not be irrotational (e.g., solid body rotation). Solutions obtained for the class of flow defined by irrotationality are thus approximations of full Navier– Stokes solutions. Mathematically, the approximation is that vorticity is negligibly small, Irrotational approximation:
→
→
→
z !§ "V !0
(10–19)
We now examine the effect of this approximation on both the continuity and momentum equations.
Irrotational flow region
Rotational flow region
FIGURE 10–22 The irrotational flow approximation is appropriate only in certain regions of the flow where the vorticity is negligible.
Continuity Equation
→
→
→
→
Thus, if § " V ! 0, then V ! § f.
(10–20)
This can easily be proven in Cartesian coordinates (Fig. 10–23), but applies to any orthogonal coordinate system as long as f is a smooth function. In words, if the curl of a vector is zero, the vector can be expressed as the gradient of a scalar function f, called the potential function. In fluid mechan→ is the velocity vector, the curl of which is the vorticity vector ics, vector V → z , and thus we call f the velocity potential function. We write For irrotational regions of flow:
→
→
V ! §f
→
∆
→
→
∆
→
§ " §f ! 0
∆
Vector identity:
Proof of the vector identity: → → " F!0 Expand in Cartesian coordinates, ∆
If you shake some more dust off your vector algebra book, you will find a vector identity concerning the curl of the gradient of any scalar function f, → and hence the curl of any vector V ,
" f!
a
∂2f ∂2f → ∂2f ∂2f → bi $a b # # ∂y ∂z ∂z ∂y ∂z ∂x ∂x ∂z j $a
∂2f ∂2f → bk ! 0 # ∂x ∂y ∂y ∂x
The identity is proven if F is a smooth function of x,, y,, and z.
(10–21)
We should point out that the sign convention in Eq. 10–21 is not universal—in some fluid mechanics textbooks, a negative sign is inserted in the definition of the velocity potential function. We state Eq. 10–21 in words as follows: In an irrotational region of flow, the velocity vector can be expressed as the gradient of a scalar function called the velocity potential function.
Regions of irrotational flow are therefore also called regions of potential flow. Note that we have not restricted ourselves to two-dimensional flows;
FIGURE 10–23 The vector identity of Eq. 10–20 is easily proven by expanding the terms in Cartesian coordinates.
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486 FLUID MECHANICS
Eq. 10–21 is valid for fully three-dimensional flow fields, as long as the approximation of irrotationality is appropriate in the region of flow under study. In Cartesian coordinates, u!
"f "x
v!
"f "y
w!
"f "z
(10–22)
and in cylindrical coordinates,
+ 2f ! 0
ur !
"f "r
uu !
1 "f r "u
uz !
"f "z
(10–23)
The usefulness of Eq. 10–21 becomes apparent→ when it is substituted into → → → Eq. 10–1, the incompressible continuity equation: § $ V ! 0 → § $ §f ! 0, or + 2f ! 0
For irrotational regions of flow:
§ 2f ! 0
(10–24) →
FIGURE 10–24 The Laplace equation for velocity potential function f is valid in both two and three dimensions and in any coordinate system, but only in irrotational regions of flow (generally away from walls and wakes).
§ 2f !
Approximation Irrotational region of flow: • Unknowns = f and P • Two equations required
FIGURE 10–25 In irrotational regions of flow, three unknown scalar components of the velocity vector are combined into one unknown scalar function—the velocity potential function.
"2f "2f "2f # # !0 "x 2 "y 2 "z 2
and in cylindrical coordinates, § 2f !
General 3-D incompressible flow: • Unknowns = u, v, w, and P • Four equations required
→
where the Laplacian operator +2 is a scalar operator defined as § $ §, and Eq. 10–24 is called the Laplace equation. We stress that Eq. 10–24 is valid only in regions where the irrotational flow approximation is reasonable (Fig. 10–24). In Cartesian coordinates,
"f 1 " 1 "2f "2f # !0 ar b# 2 r "r "r r "u 2 "z 2
The beauty of this approximation is that we have combined three unknown velocity components (u, v, and w, or ur, uu, and uz, depending on our choice of coordinate system) into one unknown scalar variable f, eliminating two of the equations required for a solution (Fig. 10–25). Once we obtain a solution of Eq. 10–24 for f, we can calculate all three components of the velocity field using Eq. 10–22 or 10–23. The Laplace equation is well known since it shows up in several fields of physics, applied mathematics, and engineering. Various solution techniques, both analytical and numerical, are available in the literature. Solutions of the Laplace equation are dominated by the geometry (i.e., boundary conditions). Although Eq. 10–24 comes from conservation of mass, mass itself (or density, which is mass per unit volume) has dropped out of the equation altogether. With a given set of boundary conditions surrounding the entire irrotational region of the flow field, we can thus solve Eq. 10–24 for f, regardless →of the fluid properties. Once we have calculated f, we can then calculate V everywhere in that region of the flow field (using Eq. 10–21), without ever having to solve the Navier–Stokes equation. The solution is valid for any incompressible fluid, regardless of its density or its viscosity, in regions of the flow in which the irrotational approximation is appropriate. The solution is even valid instantaneously for an unsteady flow, since time does not appear in the incompressible continuity equation. In other words, at any moment of time, the incompressible flow field instantly adjusts itself so as to satisfy the Laplace equation and the boundary conditions that exist at that moment of time.
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487 CHAPTER 10
Momentum Equation
We now turn our attention to the differential equation of conservation of linear momentum—the Navier–Stokes equation (Eq. 10–2). We have just shown that in an irrotational region of flow, we can obtain the velocity field without application of the Navier–Stokes equation. Why then do we need it at all? The answer is that once we have established the velocity field through use of the velocity potential function, we use the Navier–Stokes equation to solve for the pressure field. The Navier–Stokes equation is the second required equation mentioned in Fig. 10–25 for solution of two unknowns, f and P, in an irrotational region of flow. We begin our analysis by applying the irrotational flow approximation, (Eq. 10–21), to the viscous term of the Navier–Stokes equation (Eq. 10–2). Provided that f is a smooth function, that term becomes →
→
→
m § 2V ! m § 2(§ f) ! m§ (§ 2f) ! 0
V
r, m
0
where we have applied Eq. 10–24. Thus, the Navier–Stokes equation reduces to the Euler equation in irrotational regions of the flow, →
For irrotational regions of flow:
→ → → → "V → rc # (V $ § )V d ! %§ P # rg "t
We emphasize that although we get the same Euler equation as we did for an inviscid region of flow (Eq. 10–13), the viscous term vanishes here for a different reason, namely, that the flow in this region is assumed to be irrotational rather than inviscid (Fig. 10–26).
Derivation of the Bernoulli Equation in Irrotational Regions of Flow
In Section 10–4 we derived the Bernoulli equation along a streamline for inviscid regions of flow, based on the Euler equation. We now do a similar derivation beginning with Eq. 10–25 for irrotational regions of flow. For simplicity, we again assume steady incompressible flow. We use the same vector identity used previously (Eq. 10–14), leading to the alternative →form of the Euler equation of Eq. 10–15. Here, however, the vorticity vector z is negligibly small since we are considering an irrotational region of flow (Eq. 10–19). Thus, for gravity acting in the negative z-direction, Eq. 10–17 reduces to → P V2 # gzb ! 0 §a # r 2
(10–26)
We now argue that if the gradient of some scalar quantity (the quantity in parentheses in Eq. 10–26) is zero everywhere, the scalar quantity itself must be a constant. Thus, we generate the Bernoulli equation for irrotational regions of flow, Steady incompressible Bernoulli equation in irrotational regions of flow: P V2 # # gz ! C ! constant everywhere r 2
Streamlines
(10–25)
(10–27)
It is useful to compare Eqs. 10–18 and 10–27. In an inviscid region of flow, the Bernoulli equation holds along streamlines, and the Bernoulli constant
Cr
→
→ → → → ∂V + (V→ • ∇)V D = –∇P + rg→ + m∇2V ∂t
0
FIGURE 10–26 An irrotational region of flow is a region where net viscous forces are negligible compared to inertial and/or pressure forces because of the irrotational approximation. All irrotational regions of flow are therefore also inviscid, but not all inviscid regions of flow are irrotational. The fluid itself is still a viscous fluid in either case.
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488 FLUID MECHANICS Calculate f from continuity: ∇2f = 0
→
→
→
Calculate V from irrotationality: V = ∇f
Calculate P from Bernoulli: P V2 + + gz = C r 2
FIGURE 10–27 Flowchart for obtaining solutions in an irrotational region of flow. The velocity field is obtained from continuity and irrotationality, and then pressure is obtained from the Bernoulli equation. Outer region r!R
Inner region
y
r u x
P ! P∞
may change from streamline to streamline. In an irrotational region of flow, the Bernoulli constant is the same everywhere, so the Bernoulli equation holds everywhere in the irrotational region of flow, even across streamlines. Thus, the irrotational approximation is more restrictive than the inviscid approximation. A summary of the equations and solution procedure relevant to irrotational regions of flow is provided in Fig. 10–27. In a region of irrotational flow, the velocity field is obtained first by solution of the Laplace equation for velocity potential function f (Eq. 10–24), followed by application of Eq. 10–21 to obtain the velocity field. To solve the Laplace equation, we must provide boundary conditions for f everywhere along the boundary of the flow field of interest. Once the velocity field is known, we use the Bernoulli equation (Eq. 10–27) to obtain the pressure field, where the Bernoulli constant C is obtained from a boundary condition on P somewhere in the flow. Example 10–4 illustrates a situation in which the flow field consists of two separate regions—an inviscid, rotational region and an inviscid, irrotational region. EXAMPLE 10–4
A horizontal slice through a tornado (Fig. 10–28) is modeled by two distinct regions. The inner or core region (0 ' r ' R) is modeled by solid body rotation—a rotational but inviscid region of flow as discussed earlier. The outer region (r . R) is modeled as an irrotational region of flow. The flow is two→ dimensional in the r u-plane, and the components of the velocity field V ! (ur, uu) are given by
Velocity components: where v ambient pressure pressure
FIGURE 10–28 A horizontal slice through a tornado can be modeled by two regions—an inviscid but rotational inner region of flow (r ' R) and an irrotational outer region of flow (r . R).
A Two-Region Model of a Tornado
ur ! 0
vr u u ! cvR2 r
0'r'R r.R
(1)
is the magnitude of the angular velocity in the inner region. The pressure (far away from the tornado) is equal to P&. Calculate the field in a horizontal slice of the tornado for 0 ' r ' &. What is the at r ! 0? Plot the pressure and velocity fields.
SOLUTION We are to calculate the pressure field P(r) in a horizontal radial slice through a tornado for which the velocity components are approximated by Eq. 1. We are also to calculate the pressure in this horizontal slice at r ! 0. Assumptions 1 The flow is steady and incompressible. 2 Although R increases and v decreases with increasing elevation z, R and v are assumed to be constants when considering a particular horizontal slice. 3 The flow in the horizontal slice is two-dimensional in the ru-plane (no dependence on z and no w-component of velocity). 4 The effects of gravity are negligible within a particular horizontal slice (an additional hydrostatic pressure field exists in the z-direction, of course, but this does not affect the dynamics of the flow, as discussed previously). Analysis In the inner region, the Euler equation is an appropriate approximation of the Navier–Stokes equation, and the pressure field is found by integration. In Example 10–3 we showed that for solid body rotation,
Pressure field in inner region (r ' R):
P!r
v 2r 2 # P0 2
(2)
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489 CHAPTER 10
where P0 is the (unknown) pressure at r ! 0 and we have neglected the gravity term. Since the outer region is a region of irrotational flow, the Bernoulli equation is appropriate and the Bernoulli constant is the same everywhere from r ! R outward to r → &. The Bernoulli constant is found by applying the boundary condition far from the tornado, namely, as r → &, uu → 0 and P → P& (Fig. 10–29). Equation 10–27 yields
V2 P # r 2
As r → &:
gz
#
!C
→
C!
F
P& r
Hint of
y
the Da
the far Look to e you her field. T what d fi may n k. u o y see
(3)
F
F P & /r
V →0 as r→&
assumption 4
The pressure field anywhere in the outer region is obtained by substituting the value of constant C from Eq. 3 into the Bernoulli equation (Eq. 10–27). Neglecting gravity,
In outer region (r . R):
1 1 P ! rC % rV 2 ! P& % rV 2 2 2
FIGURE 10–29 A good place to obtain boundary conditions for this problem is the far field; this is true for many problems in fluid mechanics.
(4)
We note that V 2 ! uu2. After substitution of Eq. 1 for uu, Eq. 4 reduces to
P ! P% $
Pressure field in outer region (r . R):
R V2R 4 2 r2
(5)
At r ! R, the interface between the inner and outer regions, the pressure must be continuous (no sudden jumps in P), as illustrated in Fig. 10–30. Equating Eqs. 2 and 5 at this interface yields
Pressure at r ! R:
Pr!R ! r
r v 2R4 v 2R2 # P0 ! P& % 2 2 R2
P
(6) r=R
from which the pressure P0 at r ! 0 is found,
Pressure at r ! 0:
P0 ! P% $ RV2R 2
(a) (7)
Equation 7 provides the value of pressure in the middle of the tornado—the eye of the storm. This is the lowest pressure in the flow field. Substitution of Eq. 7 into Eq. 2 enables us to rewrite Eq. 2 in terms of the given far-field ambient pressure P&,
In inner region (r ' R):
r2 P ! P% $ RV2 aR 2 % b 2
(8)
Instead of plotting P as a function of r in this horizontal slice, we plot a nondimensional pressure distribution instead, so that the plot is valid for any horizontal slice. In terms of nondimensional variables,
Inner region (r ' R):
uu r ! vR R
Outer region (r . R):
uu R ! vR r
P % P& 1 r 2 ! a b %1 2 R rv 2R2 P % P& 1 R 2 ! % a b 2 r rv 2R2
(9)
Figure 10–31 shows both nondimensional tangential velocity and nondimensional pressure as functions of nondimensional radial location. Discussion In the outer region, pressure increases as speed decreases—a direct result of the Bernoulli equation, which applies with the same Bernoulli constant everywhere in the outer region. You are encouraged to calculate P
r
P
r=R (b)
r
FIGURE 10–30 For our model of the tornado to be valid, the pressure can have a discontinuity in slope at r ! R, but cannot have a sudden jump of value there; (a) is valid, but (b) is not.
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490 FLUID MECHANICS 0
1 Nondimensional pressure
–0.2
0.8 Inner region
Outer region –0.4
0.6
FIGURE 10–31 Nondimensional tangential velocity distribution (blue curve) and nondimensional pressure distribution (black curve) along a horizontal radial slice through a tornado. The inner and outer regions of flow are marked.
uu vR 0.4
P – P∞ 2 2 –0.6 rv R
0.2
–0.8
Nondimensional tangential velocity
–1
0 0
1
2
3
4
5
r/R
Auntie Em!
FIGURE 10–32 The lowest pressure occurs at the center of the tornado, and the flow in that region can be approximated by solid body rotation.
in the outer region by an alternate method—direct integration of the Euler equation without use of the Bernoulli equation; you should get the same result. In the inner region, P increases parabolically with r even though speed also increases; this is because the Bernoulli constant changes from streamline to streamline (as also pointed out in Example 10–3). Notice that even though there is a discontinuity in the slope of tangential velocity at r/R ! 1, the pressure has a fairly smooth transition between the inner and outer regions. The pressure is lowest in the center of the tornado and rises to atmospheric pressure in the far field (Fig. 10–32). Finally, the flow in the inner region is rotational but inviscid, since viscosity plays no role in that region of the flow. The flow in the outer region is irrotational and inviscid. Note, however, that viscosity still acts on fluid particles in the outer region. (Viscosity causes the fluid particles to shear and distort, even though the net viscous force on any fluid particle in the outer region is zero.)
Two-Dimensional Irrotational Regions of Flow
In irrotational regions of flow, Eqs. 10–24 and 10–21 apply for both twoand three-dimensional flow fields, and we solve for the velocity field in these regions by solving the Laplace equation for velocity potential function f. If the flow is also two-dimensional, we are able to make use of the stream function as well (Fig. 10–33). The two-dimensional approximation is not limited to flow in the xy-plane, nor is it limited to Cartesian coordinates. In fact, we can assume two-dimensionality in any region of the flow where only two directions of motion are important and where there is no significant variation in the third direction. The two most common examples are planar flow (flow in a plane with negligible variation in the direction normal to the plane) and axisymmetric flow (flow in which there is rotational symmetry about some axis). We may also choose to work in Cartesian coordinates, cylindrical coordinates, or spherical polar coordinates, depending on the geometry of the problem at hand.
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491 CHAPTER 10
Planar Irrotational Regions of Flow
We consider planar flow first, since it is the simplest. For a steady, incompressible, planar, irrotational region of flow in the xy-plane in Cartesian coordinates (Fig. 10–34), the Laplace equation for f is "2f "2f § f! 2# 2!0 "x "y 2
3-D irrotational region of flow: → → • V = ∇f 2 • ∇ f=0 • Cannot define c
(10–28)
For incompressible planar flow in the xy-plane, the stream function c is defined as (Chap. 9) Stream function:
u!
"c "y
v!%
"c "x
2-D irrotational region of flow: → → • V = ∇f 2 • ∇ f=0 • Can also use c
(10–29)
Note that Eq. 10–29 holds whether the region of flow is rotational or irrotational. In fact, the stream function is defined such that it always satisfies the continuity equation, regardless of rotationality. If we restrict our approximation to irrotational regions of flow, Eq. 10–19 must also hold; namely, the vorticity is zero or negligibly small. For general two-dimensional flow in the xy-plane, the z-component of vorticity is the only nonzero component. Thus, in an irrotational region of flow, zz !
FIGURE 10–33 Two-dimensional flow is a subset of three-dimensional flow; in twodimensional regions of flow we can define a stream function, but we cannot do so in three-dimensional flow. The velocity potential function, however, can be defined for any irrotational region of flow.
"v "u % !0 "x "y
Substitution of Eq. 10–29 into this equation yields "c "2c "2c " "c " a% b % a b !% 2% 2!0 "x "x "y "y "x "y
We recognize the Laplacian operator in this latter equation. Thus, § 2c !
"2c
"2c
"x
"y 2
# 2
!0
(10–30)
We conclude that the Laplace equation is applicable, not only for f (Eq. 10–28), but also for c (Eq. 10–30) in steady, incompressible, irrotational, planar regions of flow. Curves of constant values of c define streamlines of the flow, while curves of constant values of f define equipotential lines. (Note that some authors use the phrase equipotential lines to refer to both streamlines and lines of constant f rather than exclusively for lines of constant f.) In planar irrotational regions of flow, it turns out that streamlines intersect equipotential lines at right angles, a condition known as mutual orthogonality (Fig. 10–35). In addition, the potential functions c and f are intimately related to each other—both satisfy the Laplace equation, and from either c or f we can determine the velocity field. Mathematicians call solutions of c and f harmonic functions, and c and f are called harmonic conjugates of each other. Although c and f are related, their origins are somewhat opposite; it is perhaps best to say that c and f are complementary to each other: • The stream function is defined by continuity; the Laplace equation for c results from irrotationality. • The velocity potential is defined by irrotationality; the Laplace equation for f results from continuity.
y →
V
v →
j
x →
u
i
y x
FIGURE 10–34 Velocity components and unit vectors in Cartesian coordinates for planar twodimensional flow in the xy-plane. There is no variation normal to this plane.
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492 FLUID MECHANICS
90° Streamlines
Laplace equation, planar flow in (r, u):
Equipotential lines
FIGURE 10–35 In planar irrotational regions of flow, curves of constant f (equipotential lines) and curves of constant c (streamlines) are mutually orthogonal, meaning that they intersect at 90° angles everywhere.
y
→
V
→
→
er
eu r u
"f 1 " 1 "2f ar b# 2 !0 r "r "r r "u 2
(10–31)
The stream function c for planar flow in Cartesian coordinates is defined by Eq. 10–29, and the irrotationality condition causes c to also satisfy the Laplace equation. In cylindrical coordinates we perform a similar analysis. Recall from Chap. 9, Stream function:
ur !
1 "c r "u
uu ! %
"c "r
(10–32)
It is left as an exercise for you to show that the stream function defined by Eq. 10–32 also satisfies the Laplace equation in cylindrical coordinates for regions of two-dimensional planar irrotational flow. (Verify your results by replacing f by c in Eq. 10–31 to obtain the Laplace equation for the stream function.)
Axisymmetric Irrotational Regions of Flow
uu
x
In practice, we may perform a potential flow analysis using either c or f, and we should achieve the same results either way. However, it is often more convenient to use c, since boundary conditions on c are usually easier to specify. Planar flow in the xy-plane can also be described in cylindrical coordinates (r, u) and (ur, uu), as shown in Fig. 10–36. Again, there is no z-component of velocity, and velocity does not vary in the z-direction. In cylindrical coordinates,
ur
y x
FIGURE 10–36 Velocity components and unit vectors in cylindrical coordinates for planar flow in the ru-plane. There is no variation normal to this plane.
Axisymmetric flow is a special case of two-dimensional flow that can be described in either cylindrical coordinates or spherical polar coordinates. In cylindrical coordinates, r and z are the relevant spatial variables, and ur and uz are the nonzero velocity components (Fig. 10–37). There is no dependence on angle u since rotational symmetry is defined about the z-axis. This is a type of two-dimensional flow because there are only two independent spatial variables, r and z. (Imagine rotating the radial component r in Fig. 10–37 in the u-direction about the z-axis without changing the magnitude of r.) Because of rotational symmetry about the z-axis, the magnitudes of velocity components ur and uz remain unchanged after such a rotation. The Laplace equation for velocity potential f for the case of axisymmetric irrotational regions of flow in cylindrical coordinates is "f "2f 1 " ar b # 2 !0 r "r "r "z
In order to obtain expressions for the stream function for axisymmetric flow, we begin with the incompressible continuity equation in r- and z-coordinates, "u z 1 " (ru r) # !0 r "r "z
(10–33)
After some algebra, we define a stream function that identically satisfies Eq. 10–33, Stream function:
ur ! %
1 "c r "z
uz !
1 "c r "r
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493 CHAPTER 10
Following the same procedure as for planar flow, we generate an equation for c for axisymmetric irrotational regions of flow by forcing the vorticity to be zero. In this case, only the u-component of vorticity is relevant since the velocity vector always lies in the rz-plane. Thus, in an irrotational region of flow,
y r
x
z ur
u
"u r "u z " 1 "c " 1 "c % ! a% b% a b !0 r "z "z "r "z "r r "r
Rotational symmetry
After taking r outside the z-derivative (since r is not a function of z), we get r
"2c " 1 "c a b # 2 !0 "r r "r "z
(10–34)
Note that Eq. 10–34 is not the same as the Laplace equation for c. You cannot use the Laplace equation for the stream function in axisymmetric irrotational regions of flow (Fig. 10–38). For planar irrotational regions of flow, the Laplace equation is valid for both f and c; but for axisymmetric irrotational regions of flow, the Laplace equation is valid for f but not for c.
A direct consequence of this statement is that curves of constant c and curves of constant f in axisymmetric irrotational regions of flow are not mutually orthogonal. This is a fundamental difference between planar and axisymmetric flows. Finally, even though Eq. 10–34 is not the same as the Laplace equation, it is still a linear partial differential equation. This allows us to use the technique of superposition with either c or f when solving for the flow field in axisymmetric irrotational regions of flow. Superposition is discussed shortly.
Summary of Two-Dimensional Irrotational Regions of Flow
uz
r
Axisymmetric body z
FIGURE 10–37 Flow over an axisymmetric body in cylindrical coordinates with rotational symmetry about the z-axis. Neither the geometry nor the velocity field depend on u; and uu ! 0.
CAUTION! LAPLACE EQUATION NOT AVAILABLE FOR STREAM FUNCTION IN AXISYMMETRIC FLOW
Equations for the velocity components for both planar and axisymmetric irrotational regions of flow are summarized in Table 10–2.
TA B L E 1 0 – 2 Velocity components for steady, incompressible, irrotational, two-dimensional regions of flow in terms of velocity potential function and stream function in various coordinate systems Description and Coordinate System
Velocity Component 1
Planar; Cartesian coordinates
u!
Planar; cylindrical coordinates
ur !
Axisymmetric; cylindrical coordinates
ur !
"f "c ! "x "y
Velocity Component 2 v!
"c "f !% "y "x
"f 1 "c ! r "u "r
uu !
"c 1 "f !% r "u "r
"f 1 "c !% r "z "r
uz !
"f 1 "c ! r "r "z
FIGURE 10–38 The equation for the stream function in axisymmetric irrotational flow (Eq. 10–34) is not the Laplace equation.
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494 FLUID MECHANICS
Superposition in Irrotational Regions of Flow
#
f1
#
!
f2
!
f
FIGURE 10–39 Superposition is the process of adding two or more irrotational flow solutions together to generate a third (more complicated) solution.
Since the Laplace equation is a linear homogeneous differential equation, the linear combination of two or more solutions of the equation must also be a solution. For example, if f1 and f2 are each solutions of the Laplace equation, then Af1, (A # f1), (f1 # f2), and (Af1 # Bf2) are also solutions, where A and B are arbitrary constants. By extension, one may combine several solutions of the Laplace equation, and the combination is guaranteed to also be a solution. If a region of irrotational flow is modeled by the sum of two or more separate irrotational flow fields, e.g., a source located in a free-stream flow, one can simply add the velocity potential functions for each individual flow to describe the combined flow field. This process of adding two or more known solutions to create a third, more complicated solution is known as superposition (Fig. 10–39). For the case of two-dimensional irrotational flow regions, a similar analysis can be performed using the stream function rather than the velocity potential function. We stress that the concept of superposition is useful, but is valid only for irrotational flow fields for which the equations for f and c are linear. You must be careful to ensure that the two flow fields you wish to add vectorially are both irrotational. For example, the flow field for a jet should never be added to the flow field for an inlet or for free-stream flow, because the velocity field associated with a jet is strongly affected by viscosity, is not irrotational, and cannot be described by potential functions. It also turns out that since the potential function of the composite field is the sum of the potential functions of the individual flow fields, the velocity at any point in the composite field is the vector sum of the velocities of the individual flow fields. We prove this in Cartesian coordinates by considering a planar irrotational flow field that is the superposition of two independent planar irrotational flow fields denoted by subscripts 1 and 2. The composite velocity potential function is given by Superposition of two irrotational flow fields:
f ! f1 # f2
Using the equations for planar irrotational flow in Cartesian coordinates in Table 10–2, the x-component of velocity of the composite flow is # →
V1
#
u!
! →
V2
!
→
V
FIGURE 10–40 In the superposition of two irrotational flow solutions, the two velocity vectors at any point in the flow region add vectorially to produce the composite velocity at that point.
"f "(f1 # f2) "f1 "f2 ! ! # ! u1 # u2 "x "x "x "x
We can generate an analogous expression for v. Thus, superposition enables us to simply add the individual velocities vectorially at any location in the flow region to obtain the velocity of the composite flow field at that location (Fig. 10–40). Composite velocity field from superposition:
→
→
→
V ! V1 # V2
Elementary Planar Irrotational Flows
(10–35)
Superposition enables us to add two or more simple irrotational flow solutions to create a more complex (and hopefully more physically significant) flow field. It is therefore useful to establish a collection of elementary building block irrotational flows, with which we can construct a variety of more practical flows (Fig. 10–41). Elementary planar irrotational flows are
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495 CHAPTER 10
described in xy- and/or ru-coordinates, depending on which pair is more useful in a particular problem.
ø1 ø1ø2ø1
ø ø3 ø4 ø 5 ø4 ø4 ø2 ø2 ø4 ø4 5 ø3 ø3 ø2 ø2 ø1 ø2 ø2 ø1 ø ø2 2
Building Block 1—Uniform Stream
The simplest building block flow we can think of is a uniform stream of flow moving at constant velocity V in the x-direction (left to right). In terms of the velocity potential and stream function (Table 10–2), Uniform stream:
u!
"f "c ! !V "x "y
v!
"c "f !% !0 "y "x
By integrating the first of these with respect to x, and then differentiating the result with respect to y, we generate an expression for the velocity potential function for a uniform stream, f ! Vx # f (y)
→
v!
"f ! f ((y) ! 0 "y
→
f (y) ! constant
FIGURE 10–41 With superposition we can build up a complicated irrotational flow field by adding together elementary “building block” irrotational flow fields.
The constant is arbitrary since velocity components are always derivatives of f. We set the constant equal to zero, knowing that we can always add an arbitrary constant later on if desired. Thus, Velocity potential function for a uniform stream:
f ! Vx
y c3
(10–36)
c2
In a similar manner we generate an expression for the stream function for this elementary planar irrotational flow,
c1
Stream function for a uniform stream:
c ! Vy
(10–37)
Shown in Fig. 10–42 are several streamlines and equipotential lines for a uniform stream. Notice the mutual orthogonality. It is often convenient to express the stream function and velocity potential function in cylindrical coordinates rather than rectangular coordinates, particularly when superposing a uniform stream with some other planar irrotational flow(s). The conversion relations are obtained from the geometry of Fig. 10–36, x ! r cos u
y ! r sin u
r ! 2x 2 # y 2
(10–38)
V
c=0
x
–c1 –c2 –f2
–f1
u ! u r cos u % u u sin u
v ! u r sin u # u u cos u
f ! Vr cos u
c ! Vr sin u
(10–40)
We may modify the uniform stream so that the fluid flows uniformly at speed V at an angle of inclination a from the x-axis. For this situation, u ! V cos a and v ! V sin a as shown in Fig. 10–43. It is left as an exercise to show that the velocity potential function and stream function for a uniform stream inclined at angle a are Uniform stream inclined at angle a:
f ! V(x cos a # y sin a) c ! V(y cos a % x sin a)
f1
f2
y
(10–39)
In cylindrical coordinates, Eqs. 10–36 and 10–37 for f and c become Uniform stream:
f=0
FIGURE 10–42 Streamlines (solid) and equipotential lines (dashed) for a uniform stream in the x-direction.
From Eq. 10–38 and a bit of trigonometry, we derive relationships for u and v in terms of cylindrical coordinates, Transformation:
ø4
ø3
(10–41)
When necessary, Eq. 10–41 can easily be converted to cylindrical coordinates through use of Eq. 10–38.
c2 c1
a
V
x
c=0 –c1 –c2 –f2
–f1
f=0
f1
f2
FIGURE 10–43 Streamlines (solid) and equipotential lines (dashed) for a uniform stream inclined at angle a.
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496 FLUID MECHANICS
Building Block 2—Line Source or Line Sink
y
xy-plane
L z
x
FIGURE 10–44 Fluid emerging uniformly from a finite line segment of length L. As L approaches infinity, the flow becomes a line source, and the xy-plane is taken as normal to the axis of the source.
Our second building block flow is a line source. Imagine a line segment of length L parallel to the z-axis, along which fluid emerges and flows uniformly outward in all directions normal . to the line segment (Fig. 10–44). The total volume flow rate is equal to V. As length L approaches infinity, the flow becomes two-dimensional in planes perpendicular to the line, and the line . from which the fluid escapes is called a line source. For an infinite line, V also approaches infinity; thus, it is more convenient to consider the . volume flow rate per unit depth, V/L, called the line source strength (often given the symbol m). A line sink is the opposite of a line source; fluid flows into the line from all directions in planes normal to the axis of the. line sink. By convention, . positive V/L signifies a line source and negative V/L signifies a line sink. The simplest case occurs when the line source is located at the origin of the xy-plane, with the line itself lying along the z-axis. In the xy-plane, the line source looks like a point at the origin from which fluid is spewed outward in all directions in the plane (Fig. 10–45). At any radial distance r from the line source, the radial velocity component ur is found by applying conservation of mass. Namely, the entire volume flow rate per unit depth from the line source must pass through the circle defined by radius r. Thus, # V ! 2pru r L
y ⋅ V/L
ur r u x
FIGURE 10–45 . Line source of strength V/L located at the origin in the xy-plane; the total volume flow rate per unit depth through. a circle of any radius r must equal V/L regardless of the value of r.
# V ,L ur ! 2pr
(10–42)
Clearly, ur decreases with increasing r as we would expect. Notice also that ur is infinite at the origin since r is zero in the denominator of Eq. 10–42. We call this a singular point or a singularity—it is certainly unphysical, but keep in mind that planar irrotational flow is merely an approximation, and the line source is still useful as a building block for superposition in irrotational flow. As long as we stay away from the immediate vicinity of the center of the line source, the rest of the flow field produced by superposition of a line source and other building block(s) may still be a good representation of a region of irrotational flow in a physically realistic flow field. We now generate expressions for the velocity . potential function and the stream function for a line source of strength V/L. We use cylindrical coordinates, beginning with Eq. 10–42 for ur and also recognizing that uu is zero everywhere. Using Table 10–2, the velocity components are Line source:
ur !
# "f 1 "c V ,L ! ! r "u 2pr "r
uu !
"c 1 "f !% !0 r "u "r
To generate the stream function, we (arbitrarily) choose one of these equations (we choose the second one), integrate with respect to r, and then differentiate with respect to the other variable u, "c ! %u u ! 0 "r
→
c ! f (u)
→
# "c V ,L ! f ((u) ! ru r ! "u 2p
from which we integrate to obtain
# V ,L u # constant f (u) ! 2p
Again we set the arbitrary constant of integration equal to zero, since we can add back a constant as desired at any time without changing the flow.
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497 CHAPTER 10
After a similar analysis for f, we obtain the following expressions for a line source at the origin: Line source at the origin:
f!
# V ,L ln r 2p
and
c!
# V ,L u 2p
y c3
(10–43)
Several streamlines and equipotential lines are sketched for a line source in Fig. 10–46. As expected, the streamlines are rays (lines of constant u), and the equipotential lines are circles (lines of constant r). The streamlines and equipotential lines are mutually orthogonal everywhere except at the origin, a singularity point. In situations where we would like to place a line source somewhere other than the origin, we must transform Eq. 10–43 carefully. Sketched in Fig. 10–47 is a source located at some arbitrary point (a, b) in the xy-plane. We define r1 as the distance from the source to some point P in the flow, where P is located at (x, y) or (r, u). Similarly, we define u1 as the angle from the source to point P, as measured from a line parallel to the x-axis. We analyze the flow as if the source were at a new origin at absolute location (a, b). Equations 10–43 for f and c are thus still usable, but r and u must be replaced by r1 and u1. Some trigonometry is required to convert r1 and u1 back to (x, y) or (r, u). In Cartesian coordinates, for example,
c4
r
c2
u c5
c1
f1 f2
x
f3
c8
c6 c7
FIGURE 10–46 Streamlines (solid) and equipotential lines (dashed) for a line source of . strength V/L located at the origin in the xy-plane.
# # V ,L V ,L ln r1 ! ln 2(x % a)2 # (y % b)2 2p 2p (10–44) # # y%b V ,L V ,L u ! arctan c! x%a 2p 1 2p
y
f!
Line source at point (a, b):
⋅ V/L
P ⋅ V/L
r
r1 u1
EXAMPLE 10–5
Superposition of a Source and Sink of Equal Strength
b
Consider an irrotational region of flow composed of a line source of strength . V/L at location (%a, 0) and a line sink of the same strength (but opposite sign) at (a, 0), as sketched in Fig. 10–48. Generate an expression for the stream function in both Cartesian and cylindrical coordinates.
SOLUTION We are to superpose a source and a sink, and generate an expression for c in both Cartesian and cylindrical coordinates. Assumptions The region of flow under consideration is incompressible and irrotational. Analysis We use Eq. 10–44 to obtain c for the source, Line source at (%a, 0):
# V ,L c1 ! u1 2p
where
y u 1 ! arctan x#a
u x
a
FIGURE . 10–47 Line source of strength V/L located at some arbitrary point (a, b) in the xy-plane. y
P r1
(1)
r
Similarly for the sink,
Line sink at (a, 0):
# %V ,L c2 ! u 2p 2
where
y u 2 ! arctan x%a
(2)
Superposition enables us to simply add the two stream functions, Eqs. 1 and 2, to obtain the composite stream function,
Composite stream function:
# V ,L c ! c1 # c2 ! (u % u 2) 2p 1
u
u1
(3)
•
V/L a
•
– V/L a
r2 u2
x
FIGURE 10–48 Superposition of a line source of . strength V/L at (%a, 0) and . a line sink (source of strength %V/L) at (a, 0).
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498 FLUID MECHANICS
We rearrange Eq. 3 and take the tangent of both sides to get
Useful Trigonometric Identities sin( # b)) ! sin a cos b # cos a sin b sin(a cos(a # b)) ! cos a cos b % sin a sin b cos( tan a # tan b tan(a # b)) ! tan( 1 – tan a tan b cot b cot a – 1 cot(a # b)) ! cot( cot b # cot a
2pc tan u 1 % tan u 2 tan # ! tan (u 1 % u 2) ! 1 # tan u 1 tan u 2 V ,L
(4)
where we have used a trigonometric identity (Fig. 10–49). We substitute Eqs. 1 and 2 for u1 and u2 and perform some algebra to obtain an expression for the stream function,
y y % 2pc %2ay x#a x%a ! 2 tan # ! y y x # y 2 % a2 V ,L 1# x#ax%a
FIGURE 10–49 Some useful trigonometric identities.
or, taking the arctangent of both sides,
Final result, Cartesian coordinates:
# 2ay $V ,L C! arctan 2 2P x # y2 $ a2
(5)
We translate to cylindrical coordinates by using Eq. 10–38,
# $V ,L 2ar sin U C! arctan 2 2P r $ a2
Final result, cylindrical coordinates:
(6)
Discussion If the source and sink were to switch places, the result . would be the same, except that the negative sign on source strength V/L would disappear.
y uu r
Building Block 3—Line Vortex
u
L
x
FIGURE 10–50 Line vortex of strength / located at the origin in the xy-plane. y
L b
Line vortex:
ur !
"f 1 "c ! !0 r "u "r
r1 u1
Line vortex at the origin: u a
FIGURE 10–51 Line vortex of strength / located at some arbitrary point (a, b) in the xy-plane.
uu !
"c / 1 "f !% ! r "u "r 2pr
(10–45)
where / is called the circulation or the vortex strength. Following the standard convention in mathematics, positive / represents a counterclockwise vortex, while negative / represents a clockwise vortex. It is left as an exercise to integrate Eq. 10–45 to obtain expressions for the stream function and the velocity potential function,
P r
Our third building block flow is a line vortex parallel to the z-axis. As with the previous building block, we start with the simple case in which the line vortex is located at the origin (Fig. 10–50). Again we use cylindrical coordinates for convenience. The velocity components are
x
f!
/ u 2p
/ c ! % ln r 2p
(10–46)
Comparing Eqs. 10–43 and 10–46, we see that a line source and line vortex are somewhat complementary in the sense that the expressions for f and c are reversed. In situations where we would like to place the vortex somewhere other than the origin, we must transform Eq. 10–46 as we did for a line source. Sketched in Fig. 10–51 is a line vortex located at some arbitrary point (a, b)
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499 CHAPTER 10
in the xy-plane. We define r1 and u1 as previously (Fig. 10–47). To obtain expressions for f and c, we replace r and u by r1 and u1 in Eqs. 10–46 and then transform to regular coordinates, either Cartesian or cylindrical. In Cartesian coordinates, f! Line vortex at point (a, b):
EXAMPLE 10–6
y%b / / u1 ! arctan x%a 2p 2p
(10–47)
/ / ln r1 ! % ln 2(x % a)2 # (y % b)2 c!% 2p 2p
Velocity in a Flow Composed of Three Components
An . irrotational region of flow is formed by superposing a line source of strength . (V/L)1 ! 2.00 m2/s at (x, y) ! (0, %1), a line source of strength (V/L)2 ! %1.00 m2/s at (x, y) ! (1, %1), and a line vortex of strength / ! 1.50 m2/s at (x, y) ! (1, 1), where all spatial coordinates are in meters. . [Source number 2 is actually a sink, since (V/L)2 is negative.] The locations of the three building blocks are shown in Fig. 10–52. Calculate the fluid velocity at the point (x, y) ! (1, 0).
y, m
SOLUTION For the given superposition of two line sources and a vortex, we
0
are to calculate the velocity at the point (x, y) ! (1, 0). Assumptions 1 The region of flow being modeled is steady, incompressible, and irrotational. 2 The velocity at the location of each component is infinite (they are singularities), and the flow in the vicinity of each of these singularities is unphysical; however, these regions are ignored in the present analysis. Analysis There are several ways to solve this problem. We could sum the three stream functions using Eqs. 10–44 and 10–47, and then take derivatives of the composite stream function to calculate the velocity components. Alternatively, we could do the same for velocity potential function. An easier approach is to recognize that velocity itself can be superposed; we simply add the velocity vectors induced by each of the three individual singularities to form the composite velocity at the given point. This is illustrated in Fig. 10–53. Since the vortex is located 1 m above the point (1, 0), the velocity induced by the vortex is to the right and has a magnitude of
Vvortex !
/ 1.50 m2,s ! ! 0.239 m,s 2prvortex 2p(1.00 m)
(1)
Similarly, the first source induces a velocity at point (1, 0) at a 45° angle from the x-axis as shown in Fig. 10–53. Its magnitude is
Vsource 1 !
# 0 (V ,L)1 0
2prsource 1
!
2.00 m2/s 2p( 22 m)
! 0.225 m/s
(2)
Finally, the second source (the sink) induces a velocity straight down with magnitude
Vsource 2 !
# 0 (V ,L)2 0
2prsource 2
!
0 %1.00 m2,s 0 2p(1.00 m)
! 0.159 m,s
(3)
L
1
rvortex Point of interest 0
1 x, m rsource 1
•
(V/L)1 –1
rsource 2 •
(V/L)2
FIGURE 10–52 Superposition of two line sources and a line vortex in the xy-plane (Example 10–6).
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500 FLUID MECHANICS y, m
We sum these velocities vectorially by completing the parallelograms, as illustrated in Fig. 10–54. Using Eq. 10–35, the resultant velocity is
L
1
→
→
0
0
→
0.239i m/s →
Vvortex
1
x, m (a)
→
Vsource 1
y, m 0
0
→
→
(4)
→
%0.159j m/s
The superposed velocity at point (1, 0) is 0.398 m/s to the right. Discussion This example demonstrates that velocity can be superposed just as stream function or velocity potential function can be superposed. Superposition of velocity is valid in irrotational regions of flow because the differential equations for f and c are linear; the linearity extends to their derivatives as well.
1 x, m
Building Block 4—Doublet
rsource 1
Our fourth and final building block flow is called a doublet. Although we treat it as a building block for use with superposition, the doublet itself is generated by superposition of two earlier building blocks, namely, a line source and a line sink of equal magnitude, as discussed in Example 10–5. The composite stream function was obtained in that example problem and the result is repeated here:
•
(V/L)1
–1
(b) y, m 0
0.225 → 0.225 → a i # j b m,s 22 22
→
Vsource 2 ! (0.398i # 0 j ) m,s
rvortex
→
Vsource 1 #
V ! Vvortex #
0
Composite stream function:
1 x, m
→
Vsource 2
•
(V/L)2
Stream function as a → 0: (c)
FIGURE 10–53 Induced velocity due to (a) the vortex, (b) source 1, and (c) source 2 (noting that source 2 is negative) (Example 10–6). →
→
Vvortex
x
Vsource 2 Resultant velocity
FIGURE 10–54 Vector summation of the three induced velocities of Example 10–6.
# %a(V ,L)r sin u c→ p(r 2 % a 2)
(10–49)
. If we shrink a while maintaining the same source and sink strengths (V/L . and %V/L), the source and sink cancel each other out when a ! 0, leaving us with no flow at all. However, .imagine that as the source and sink approach each other, their. strength V/L increases inversely with distance a such that the product a(V/L) remains constant. In that case, r .. a at any point P except very close to the origin, and Eq. 10–49 reduces to Doublet along the x-axis:
Vsource 1 Point (1, 0)
→
(10–48)
Now imagine that the distance a from the origin to the source and from the origin to the sink approaches zero (Fig. 10–55). You should recall that arctan b approaches b for very small values of b. Thus, as distance a approaches zero, Eq. 10–48 reduces to
rsource 2
–1
# %V ,L 2ar sin u c! arctan 2 2p r % a2
# %a(V ,L) sin u sin u c! ! %K p r r
(10–50)
. where we have defined doublet strength K ! a(V/L)/p for convenience. The velocity potential function is obtained in similar fashion, Doublet along the x-axis:
f!K
cos u r
(10–51)
Several streamlines and equipotential lines for a doublet are plotted in Fig. 10–56. It turns out that the streamlines are circles tangent to the x-axis, and the equipotential lines are circles tangent to the y-axis. The circles intersect at 90° angles everywhere except at the origin, which is a singularity point.
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501 CHAPTER 10
If K is negative, the doublet is “backwards,” with the sink located at x ! 0% (infinitesimally to the left of the origin) and the source located at x ! 0# (infinitesimally to the right of the origin). In that case all the streamlines in Fig. 10–56 would be identical in shape, but the flow would be in the opposite direction. It is left as an exercise to construct expressions for a doublet that is aligned at some angle a from the x-axis.
y
P r1 r
•
V /L → ∞
r2 u2
u
u1
x
•
– V /L → –∞
Irrotational Flows Formed by Superposition
Now that we have a set of building block irrotational flows, we are ready to construct some more interesting irrotational flow fields by the superposition technique. We limit our examples to planar flows in the xy-plane; examples of superposition with axisymmetric flows can be found in more advanced textbooks (e.g., Kundu, 1990; Panton, 1996; Heinsohn and Cimbala, 2003). Note that even though c for axisymmetric irrotational flow does not satisfy the Laplace equation, the differential equation for c (Eq. 10–34) is still linear, and thus superposition is still valid.
a→ 0
a→ 0
FIGURE 10–55 A doublet is formed by superposition of a line source at (%a, 0) and a line sink . at (a, 0); a decreases to zero while V /L increases to. infinity such that the product aV /L remains constant.
Superposition of a Line Sink and a Line Vortex
. . Our first example is superposition of a line source of strength V/L (V/L is a negative quantity in this example) and a line vortex of strength /, both located at the origin (Fig. 10–57). This represents a region of flow above a drain in a sink or bathtub where fluid spirals in toward the drain. We can superpose either c or f. We choose c and generate the composite stream function by adding c for a source (Eq. 10–43) and c for a line vortex (Eq. 10–46), Superposition:
# / V ,L u% ln r c! 2p 2p
# (V ,L)u % 2pc b /
r ! exp a
ur !
# 1 "c V ,L ! r "u 2pr
r
c1
K
(10–52)
(10–53)
. We pick some arbitrary values for V/L and / so that we can generate . a plot; . 2/s. Note that V/L is negm namely, we set V/L ! %1.00 m2/s and / ! 1.50 . ative for a sink. Also note that the units for V/L and / can be obtained easily since we know that the dimensions of stream function in planar flow are {length2/time}. Streamlines are calculated for several values of c using Eq. 10–53 and are plotted in Fig. 10–58. The velocity components at any point in the irrotational region of flow are obtained by differentiating Eq. 10–52, Velocity components:
c3 c2
u
To plot streamlines of the flow, we pick a value of c and then solve for either r as a function of u or u as a function of r. We choose the former; after some algebra we get Streamlines:
y
uu ! %
"c / ! "r 2pr
We notice that in this simple example, the radial velocity component is due entirely to the sink, since there is no contribution to radial velocity from the vortex. Similarly, the tangential velocity component is due entirely to the vortex. The composite velocity at any point in the flow is the vector sum of these two components, as sketched in Fig. 10–57.
–f1
x
f1
–f3 –f2
f2
f3
–c1 –c2 –c3
FIGURE 10–56 Streamlines (solid) and equipotential lines (dashed) for a doublet of strength K located at the origin in the xy-plane and aligned with the x-axis.
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Superposition of a Uniform Stream and a Doublet— Flow over a Circular Cylinder
→
y
Vvortex →
Our next example is a classic in the field of fluid mechanics, namely, the superposition of a uniform stream of speed V& and a doublet of strength K located at the origin (Fig. 10–59). We superpose the stream function by adding Eq. 10–37 for a uniform stream and Eq. 10–50 for a doublet at the origin. The composite stream function is thus
V
→
Vsink
L
x •
•
V/L (V/L is negative here.)
c ! V& r sin u % K
Superposition:
FIGURE 10–57 Superposition of a line source of . strength V/L and a line vortex of strength / located at the origin. Vector velocity addition is shown at some arbitrary location in the xy-plane. 0.6 0.5
K ! V&a 2
Doublet strength:
and Eq. 10–54 becomes c ! V& sin uar %
0.8 0.4 0.9
c* !
0.3
–1
0.2
0.1
–1
0 x, m
1
2
FIGURE 10–58 Streamlines created by superposition of a line sink and a line vortex at the origin. Values of c are in units of m2/s. y →
→
V∞
Vuniform stream
Vdoublet K
(10–55)
c V&a
r* !
r a
u
where angle u is already dimensionless. In terms of these parameters, Eq. 10–55 is written as
0
–2 –2
a2 b r
It is clear from Eq. 10–55 that one of the streamlines (c ! 0) is a circle of radius a (Fig. 10–60). We can plot this and other streamlines by solving Eq. 10–55 for r as a function of u or vice versa. However, as you should be aware by now, it is usually better to present results in terms of nondimensional parameters. By inspection, we define three nondimensional parameters,
0.7
1 y, m 0
(10–54)
For convenience we set c ! 0 when r ! a (the reason for this will soon become apparent). Equation 10–54 can then be solved for doublet strength K,
Alternate form of stream function: 2
sin u r
→
V
x
FIGURE 10–59 Superposition of a uniform stream and a doublet; vector velocity addition is shown at some arbitrary location in the xy-plane.
c* ! sin u ar* %
1 b r*
(10–56)
We solve Eq. 10–56 for r* as a function of u through use of the quadratic rule, Nondimensional streamlines:
r* !
c* 0 2(c*)2 # 4 sin2 u 2 sin u
(10–57)
Using Eq. 10–57, we plot several nondimensional streamlines in Fig. 10–61. Now you see why we chose the circle r ! a (or r* ! 1) as the zero streamline—this streamline can be thought of as a solid wall, and this flow represents potential flow over a circular cylinder. Not shown are streamlines inside the circle—they exist, but are of no concern to us. There are two stagnation points in this flow field, one at the nose of the cylinder and one at the tail. Streamlines near the stagnation points are far apart since the flow is very slow there. By contrast, streamlines near the top and bottom of the cylinder are close together, indicating regions of fast flow. Physically, fluid must accelerate around the cylinder since it is acting as an obstruction to the flow. Notice also that the flow is symmetric about both the x- and y-axes. While top-to-bottom symmetry is not surprising, fore-to-aft symmetry is perhaps unexpected, since we know that real flow around a cylinder generates a
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503 CHAPTER 10
wake region behind the cylinder, and the streamlines are not symmetric. However, we must keep in mind that the results here are only approximations of a real flow. We have assumed irrotationality everywhere in the flow field, and we know that this approximation is not true near walls and in wake regions. We calculate the velocity components everywhere in the flow field by differentiating Eq. 10–55, 2
ur !
1 "c a ! V& cos ua1 % 2 b r "u r
y V∞ K x r=a
2
"c a u u ! % ! %V& sin ua1 # 2 b "r r
(10–58)
FIGURE 10–60 Superposition of a uniform stream and a doublet yields a streamline that is a circle.
A special case is on the surface of the cylinder itself (r ! a), where Eqs. 10–58 reduce to On the surface of the cylinder:
c=0
ur ! 0
u u ! %2V& sin u
(10–59) 2
Since the no-slip condition at solid walls cannot be satisfied when making the irrotational approximation, there is slip at the cylinder wall. In fact, at the top of the cylinder (u ! 90°), the fluid speed at the wall is twice that of the free stream.
1 y*
1
0 c* = 0
EXAMPLE 10–7
Pressure Distribution on a Circular Cylinder
–1
Using the irrotational flow approximation, calculate and plot the nondimensional static pressure distribution on the surface of a circular cylinder of radius a in a uniform stream of speed V& (Fig. 10–62). Discuss the results. The pressure far away from the cylinder is P&.
SOLUTION We are to calculate and plot the nondimensional static pressure distribution along the surface of a circular cylinder in a free-stream flow. Assumptions 1 The region of flow being modeled is steady, incompressible, and irrotational. 2 The flow field is two-dimensional in the xy-plane. Analysis First of all, static pressure is the pressure that would be measured by a pressure probe moving with the fluid. Experimentally, we measure this pressure on a surface through use of a static pressure tap, which is basically a tiny hole drilled normal to the surface (Fig. 10–63). At the other end of the tap is a pressure measuring device. Experimental data of the static pressure distribution along the surface of a cylinder are available in the literature, and we compare our results to some of those experimental data. From Chap. 7 we recognize that the appropriate nondimensional pressure is the pressure coefficient, Pressure coefficient:
P%P Cp ! 1 2 & 2 rV &
V2 P V2 P # ! constant ! & # & r r 2 2
–2
–1
0 x*
1
2
FIGURE 10–61 Nondimensional streamlines created by superposition of a uniform stream and a doublet at the origin; c* ! c/(V&a), 1c* ! 0.2, x* ! x/a, and y* ! y/a, where a is the cylinder radius. y V∞ b
a
u x
(1)
Since the flow in the region of interest is irrotational, we use the Bernoulli equation (Eq. 10–27) to calculate the pressure anywhere in the flow field. Ignoring the effects of gravity,
Bernoulli equation:
–2
(2)
b=p–u
FIGURE 10–62 Planar flow over a circular cylinder of radius a immersed in a uniform stream of speed V& in the xy-plane. Angle b is defined from the front of the cylinder by convention.
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504 FLUID MECHANICS Pressure tap
Rearranging Eq. 2 into the form of Eq. 1, we get
P
Cp ! Body surface
FIGURE 10–63 Static pressure on a surface is measured through use of a static pressure tap connected to a pressure manometer or electronic pressure transducer.
Front SP
Cp in terms of angle b:
Rear SP
Free-stream pressure 0
–1
–2 Top –3 0
30
60
90 120 b, degrees
150
180
FIGURE 10–64 Pressure coefficient as a function of angle b along the surface of a circular cylinder; the solid blue curve is the irrotational flow approximation, blue circles are from experimental data at Re ! 2 * 105 % laminar boundary layer separation, and gray circles are from typical experimental data at Re ! 7 * 105 % turbulent boundary layer separation. Data from Kundu, 1990.
Surface pressure coefficient:
Cp ! 1 %
(%2V& sin u)2 ! 1 % 4 sin2 u V 2&
In terms of angle b, defined from the front of the body (Fig. 10–62), we use the transformation b ! p % u to obtain
1
Cp
(3)
We substitute our expression for tangential velocity on the cylinder surface, Eq. 10–59, since along the surface V 2 ! uu2; Eq. 3 becomes
To pressure transducer
Flexible tubing
P % P& V2 !1% 2 1 2 V& 2 rV &
Cp ! 1 $ 4 sin2 B
(4)
We plot the pressure coefficient on the top half of the cylinder as a function of angle b in Fig. 10–64, solid blue curve. (Because of top–bottom symmetry, there is no need to also plot the pressure distribution on the bottom half of the cylinder.) The first thing we notice is that the pressure distribution is symmetric fore and aft. This is not surprising since we already know that the streamlines are also symmetric fore and aft (Fig. 10–61). The front and rear stagnation points (at b ! 0° and 180°, respectively) are marked on Fig. 10–64. The pressure coefficient is unity there, and these two points have the highest pressure in the entire flow field. In physical variables, static pressure P at the stagnation points is equal to P& # rV 2& /2. In other words, the full dynamic pressure (also called impact pressure) of the oncoming fluid is felt as a static pressure on the nose of the body as the fluid is decelerated to zero speed at the stagnation point. At the very top of the cylinder (b ! 90°), the speed along the surface is twice the free-stream velocity (V ! 2V&), and the pressure coefficient is lowest there (Cp ! %3). Also marked on Fig. 10–64 are the two locations where Cp ! 0, namely at b ! 30° and 150°. At these locations, the static pressure along the surface is equal to that of the free stream (P ! P&). Discussion Typical experimental data for laminar and turbulent flow over the surface of a circular cylinder are indicated by the blue circles and gray circles, respectively, in Fig. 10–64. It is clear that near the front of the cylinder, the irrotational flow approximation is excellent. However, for b greater than about 60°, and especially near the rear portion of the cylinder (right side of the plot), the irrotational flow results do not match well at all with experimental data. In fact, it turns out that for flow over bluff body shapes like this, the irrotational flow approximation usually does a fairly good job on the front half of the body, but a very poor job on the rear half of the body. The irrotational flow approximation agrees better with experimental turbulent data than with experimental laminar data; this is because flow separation occurs farther downstream for the case with a turbulent boundary layer, as discussed in more detail in Section 10–6.
One immediate consequence of the symmetry of the pressure distribution in Fig. 10–64 is that there is no net pressure drag on the cylinder (pressure forces in the front half of the body are exactly balanced by those on the rear half of the body). In this irrotational flow approximation, the pressure fully recovers at the rear stagnation point, so that the pressure there is the same as that at the front stagnation point. We also predict that there is no net viscous
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505 CHAPTER 10
drag on the body, since we cannot satisfy the no-slip condition on the body surface when we make the irrotational approximation. Hence, the net aerodynamic drag on the cylinder in irrotational flow is identically zero. This is one example of a more general statement that applies to bodies of any shape (even unsymmetrical shapes) when the irrotational flow approximation is made, namely, the famous paradox first stated by Jean-le-Rond d’Alembert (1717–1783) in the year 1752: D’Alembert’s paradox: With the irrotational flow approximation, the aerodynamic drag force on any nonlifting body of any shape immersed in a uniform stream is zero.
D’Alembert recognized the paradox of his statement, of course, knowing that there is aerodynamic drag on real bodies immersed in real fluids. In a real flow, the pressure on the back surface of the body is significantly less than that on the front surface, leading to a nonzero pressure drag on the body. This pressure difference is enhanced if the body is bluff and there is flow separation, as sketched in Fig. 10–65. Even for streamlined bodies, however (such as airplane wings at low angles of attack), the pressure near the back of the body never fully recovers. In addition, the no-slip condition on the body surface leads to a nonzero viscous drag as well. Thus, the irrotational flow approximation falls short in its prediction of aerodynamic drag for two reasons: it predicts no pressure drag and it predicts no viscous drag. The pressure distribution at the front end of any rounded body shape is qualitatively similar to that plotted in Fig. 10–64. Namely, the pressure at the front stagnation point (SP) is the highest pressure on the body: PSP ! P& # rV 2/2, where V is the free-stream velocity (we have dropped the subscript &), and Cp ! 1 there. Moving downstream along the body surface, pressure drops to some minimum value for which P is less than P& (Cp ' 0). This point, where the velocity just above the body surface is largest and the pressure is smallest, is often called the aerodynamic shoulder of the body. Beyond the shoulder, the pressure slowly rises. With the irrotational flow approximation, the pressure always rises back to the dynamic pressure at the rear stagnation point, where Cp ! 1. However, in a real flow, the pressure never fully recovers, leading to pressure drag as discussed previously. Somewhere between the front stagnation point and the aerodynamic shoulder is a point on the body surface where the speed just above the body is equal to V, the pressure P is equal to P&, and Cp ! 0. This point is called the zero pressure point, where the phrase is obviously based on gage pressure, not absolute pressure. At this point, the pressure acting normal to the body surface is the same (P ! P&), regardless of how fast the body moves through the fluid. This fact is a factor in the location of fish eyes (Fig. 10–66). If a fish’s eye were located closer to its nose, the eye would experience an increase in water pressure as the fish swims—the faster it would swim, the higher the water pressure on its eye would be. This would cause the soft eyeball to distort, affecting the fish’s vision. Likewise, if the eye were located farther back, near the aerodynamic shoulder, the eye would experience a relative suction pressure when the fish would swim, again distorting its eyeball and blurring its vision. Experiments have revealed that the fish’s eye is instead located very close to the zero-pressure point where P ! P&, and the fish can swim at any speed without distorting its vision. Incidentally, the back of the
Irrotational flow approximation
V
Aerodynamic drag = 0 (a) Real (rotational) flow field
V
→
FD
Aerodynamic drag ≠ 0 (b)
FIGURE 10–65 (a) D’Alembert’s paradox is that the aerodynamic drag on any nonlifting body of any shape is predicted to be zero when the irrotational flow approximation is invoked; (b) in real flows there is a nonzero drag on bodies immersed in a uniform stream.
Cp 1.0 0.5 0.0 –0.5
FIGURE 10–66 A fish’s body is designed such that its eye is located near the zero-pressure point so that its vision is not distorted while it swims. Data shown are along the side of a bluefish. Adapted from American Scientist, vol. 76, p. 32, 1988.
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506 FLUID MECHANICS
gills is located near the aerodynamic shoulder so that the suction pressure there helps the fish to “exhale.” The heart is also located near this lowest-pressure point to increase the heart’s stroke volume during rapid swimming. If we think about the irrotational flow approximation a little more closely, we realize that the circle we modeled as a solid cylinder in Example 10–7 is not really a solid wall at all—it is just a streamline in the flow field that we are modeling as a solid wall. The particular streamline we model as a solid wall just happens to be a circle. We could have just as easily picked some other streamline in the flow to model as a solid wall. Since flow cannot cross a streamline by definition, and since we cannot satisfy the no-slip condition at a wall, we state the following:
4
3
y* 2
1 c* = 0.2 1
0 –2
–1
0 x*
1
With the irrotational flow approximation, any streamline can be thought of as a solid wall.
2
FIGURE 10–67 The same nondimensionalized streamlines as in Fig. 10–61, except streamline c* ! 0.2 is modeled as a solid wall. This flow represents flow of air over a symmetric hill. •
V y L w
b
Floor
z
x
(a) y Sink
For example, we can model any streamline in Fig. 10–61 as a solid wall. Let’s take the first streamline above the circle, and model it as a wall. (This streamline has a nondimensional value of c* ! 0.2.) Several streamlines are plotted in Fig. 10–67; we have not shown any streamlines below the streamline c* ! 0.2—they are still there, it’s just that we are no longer concerned with them. What kind of flow does this represent? Well, imagine wind flowing over a hill; the irrotational approximation shown in Fig. 10–67 is representative of this flow. We might expect inconsistencies very close to the ground, and perhaps on the downstream side of the hill, but the approximation is probably very good on the front side of the hill. You may have noticed a problem with this kind of superposition. Namely, we perform the superposition first, and then try to define some physical problems that might be modeled by the flow we generate. While useful as a learning tool, this technique is not always practical in real-life engineering. For example, it is unlikely that we will encounter a hill shaped exactly like the one modeled in Fig. 10–67. Instead, we usually already have a geometry and wish to model flow over or through this geometry. There are more sophisticated superposition techniques available that are better suited to engineering design and analysis. Namely, there are techniques in which numerous sources and sinks are placed at appropriate locations so as to model flow over a predetermined geometry. These techniques can even be extended to fully three-dimensional irrotational flow fields, but require a computer because of the amount of calculations involved (Kundu, 1990). We do not discuss these techniques here. EXAMPLE 10–8
b Floor
x
(b)
FIGURE 10–68 Vacuum cleaner hose with floor attachment; (a) three-dimensional view with floor in the xz-plane, and (b) view of a slice in the xy-plane with suction modeled by a line sink.
Flow into a Vacuum Cleaner Attachment
Consider the flow of air into the floor attachment nozzle of a typical household vacuum cleaner (Fig. 10–68a). The width of the nozzle inlet slot is w ! 2.0 mm, and its length is L ! 35.0 cm. The slot is held a distance b ! 2.0 cm above . the floor, as shown. The total volume flow rate through the vacuum hose is V ! 0.110 m3/s. Predict the flow field in the center plane of the attachment (the xy-plane in Fig. 10–68a). Specifically, plot several streamlines and calculate the velocity and pressure distribution along the xaxis. What is the maximum speed along the floor, and where does it occur? Where along the floor is the vacuum cleaner most effective?
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507 CHAPTER 10
SOLUTION We are to predict the flow field in the center plane of a vacuum cleaner attachment, plot velocity and pressure along the floor (x-axis), predict the location and value of the maximum velocity along the floor, and predict where along the floor the vacuum cleaner is most effective. Assumptions 1 The flow is steady and incompressible. 2 The flow in the xyplane is two-dimensional (planar). 3 The majority of the flow field is irrotational. 4 The room is infinitely large and free of air currents that might influence the flow. Analysis We approximate the slot on the vacuum cleaner attachment as a line sink (a line source with negative source strength), located a distance b above the x-axis, as sketched in Fig. 10–68b. With this approximation, we are ignoring the finite width of the slot (w); instead we model flow into the slot as flow into the line sink, which is simply a point in the xy-plane at (0, b). We are also ignoring any effects of the hose or the body of the attachment. The strength of the line source is obtained by dividing total volume flow rate by the length L of the slot,
# V %0.110 m3,s ! ! %0.314 m2,s L 0.35 m
Strength of line source:
(1)
where we include a negative sign since this is a sink instead of a source. Clearly this line sink by itself (Fig. 10–68b) is not sufficient to model the flow, since air would flow into the sink from all directions, including up through the floor. To avoid this problem, we add another elementary irrotational flow (building block) to model the effect of the floor. A clever way to do this is through the method of images. With this technique, we place a second identical sink below the floor at point (0, %b). We call this second sink the image sink. Since the x-axis is now a line of symmetry, the x-axis is itself a streamline of the flow, and hence can be thought of as the floor. The irrotational flow . field to be analyzed is sketched in Fig. 10–69. Two sources of strength V/L are shown. The top one is called the flow source, and represents suction into the vacuum cleaner attachment.. The bottom one is the image source. Keep in mind that source strength V/L is negative in this problem (Eq. 1), so that both sources are actually sinks. We use superposition to generate the stream function for the irrotational approximation of this flow field. The algebra here is similar to that of Example 10–5; in that case we had a source and a sink on the x-axis, while here we have two sources on the y-axis. We use Eq. 10–44 to obtain c for the flow source,
Line source at (0, b):
c1 !
# V ,L u 2p 1
where u 1 ! arctan
y%b x
(2)
where u 2 ! arctan
y#b x
(3)
Similarly for the image source,
Line source at (0, %b):
# V ,L u c2 ! 2p 2
Superposition enables us to simply add the two stream functions, Eqs. 2 and 3, to obtain the composite stream function,
Composite stream function:
# V ,L c ! c1 # c2 ! (u # u 2) 2p 1
(4)
y
P r1
Flow source
u1 r •
V /L b
r2
Floor
u x b •
V /L
u2
Image source
FIGURE 10–69 Superposition of a line source of . strength V/L at (0, b) and a line source of the same strength at (0, %b). The bottom source is a mirror image of the top source, making the x-axis a streamline.
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508 FLUID MECHANICS y
We rearrange Eq. 4 and take the tangent of both sides to get c* = p
2pc tan u 1 # tan u 2 tan # ! tan(u 1 # u 2) ! 1 % tan u 1tan u 2 V ,L
•
V/L
where we have again used a trigonometric identity (Fig. 10–49). We substitute Eqs. 2 and 3 for u1 and u2 and perform some algebra to obtain our final expression for the stream function in Cartesian coordinates,
c* = 0
c* = 2p
(5)
x
c!
# 2xy V ,L arctan 2 2p x % y 2 # b2
(6)
•
V/L
We translate to cylindrical coordinates using Eq. 10–38 and nondimensionalize. After some algebra,
Nondimensional stream function:
FIGURE 10–70 The x-axis is the dividing streamline that separates air produced by the top source (blue) from air produced by the bottom source (gray).
c* = –p
4
sin 2 u cos 2 u # 1,r*2
(7)
. where c* ! 2pc/(V/L), r* ! r/b, and we used trigonometric identities from Fig. 10–49. Because of symmetry about the x-axis, all the air that is produced by the upper line source must remain above the x-axis. Likewise, all the image air that is produced at the lower line source must remain below the x-axis. If we were to color air from the upper (north) source blue, and air from the lower (south) source gray (Fig. 10–70), all the blue air would stay above the x-axis, and all the gray air would stay below the x-axis. Thus, the x-axis acts as a dividing streamline, separating the blue from the gray. Furthermore, recall from Chap. 9 that the difference in value of c from one streamline to the next in planar flow is equal to the volume flow rate per unit width flowing between the two streamlines. We set c equal to zero along the positive xaxis. Following the left-side convention, introduced in Chap. 9, we know that c on the negative x-axis must equal the total volume flow rate per unit width . produced by the upper line source, i.e., V/L. Namely,
# c%x-axis % c#x-axis ! V ,L v
3
c* ! arctan
→
c*%x-axis ! 2p
(8)
0
y* 2
1
0 –2
–1 c* = –2p
0 x*
1
2
c* = 0
FIGURE 10–71 Nondimensional streamlines for the two sources of Fig. 10–69 for the case in which the source strengths are negative (they are sinks). c* is incremented uniformly from %2p (negative x-axis) to 0 (positive x-axis), and only the upper half of the flow is shown. The flow is toward the sink at location (0, 1).
These streamlines are labeled in Fig. 10–70. In addition, the nondimensional streamline c* ! p is also labeled. It coincides with the y-axis since there is symmetry about that axis as well. The origin (0, 0) is a stagnation point, since the velocity induced by the lower source exactly cancels out that induced by the upper source. For the case of the vacuum cleaner being modeled here, the source strengths are negative (they are sinks). Thus, the direction of flow is reversed, and the values of c* are of opposite sign to those in Fig. 10–70. Using the left-side convention again, we plot the nondimensional stream function for %2p ' c* ' 0 (Fig. 10–71). To do so, we solve Eq. 7 for r* as a function of u for various values of c*,
Nondimensional streamlines:
tan c* r* ! 0 B sin 2u % cos 2 u tan c*
(9)
Only the upper half is plotted, since the lower half is symmetric .and is merely the mirror image of the upper half. For the case of negative V/L, air gets sucked into the vacuum cleaner from all directions as indicated by the arrows on the streamlines. To calculate the velocity distribution on the floor (the x-axis), we can either differentiate Eq. 6 and apply the definition of stream function for planar flow
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509 CHAPTER 10
(Eq. 10–29), or we can do a vector summation. The latter is simpler and is illustrated in Fig. 10–72 for an arbitrary location along the x-axis. The . induced velocity from the upper source (or sink) has magnitude (V/L)/(2pr1), and its direction is in line with r1 as shown. Because of symmetry, the induced velocity from the image source has identical magnitude, but its direction is in line with r2. The vector sum of these two induced velocities lies along the x-axis since the two horizontal components add together, but the two vertical components cancel each other out. After a bit of trigonometry, we conclude that
# (V ,L)x u!V! p(x 2 # b 2)
Axial velocity along the x-axis:
V2 P V2 P # ! constant ! & # & r r 2 2
(11)
V 0
To generate a pressure coefficient, we need a reference velocity for the denominator. Having none, we generate one from the known parameters, . namely Vref ! %(V /L)/b, where we insert the negative sign to make Vref posi. tive (since V/L is negative for our model of the vacuum cleaner). Then we define Cp as
Cp !
Pressure coefficient:
P % P& V2 b 2V 2 ! % ! % # 1 2 V 2ref (V ,L)2 2 rV ref
(12)
where we have also applied Eq. 11. Substituting Eq. 10 for V, we get
b 2x 2 Cp ! % 2 2 p (x # b 2)2
(13)
We introduce nondimensional variables for axial velocity and distance,
Nondimensional variables:
u* !
u ub !% # Vref V ,L
x* !
x b
(14)
We note that Cp is already nondimensional. In dimensionless form, Eqs. 10 and 13 become
Along the floor:
1 x* u* ! % p 1 # x*2
Suction source •
b
2 1 x* Cp ! % a b ! %u*2 2 p 1 # x*
(15)
Curves showing u* and Cp as functions of x* are plotted in Fig. 10–73. We see from Fig. 10–73 that u* increases slowly from 0 at x* ! %& to a maximum value of about 0.159 at x* ! %1. The velocity is positive (to the right) for negative values of x* as expected since air is being sucked into the vacuum cleaner. As speed increases, pressure decreases; Cp is 0 at x ! %& and decreases to its minimum value of about %0.0253 at x* ! %1. Between x* ! %1 and x* ! 0 the speed decreases to zero while the pressure increases to zero at the stagnation point directly below the vacuum cleaner nozzle. To the right of the nozzle (positive values of x*), the velocity is antisymmetric, while the pressure is symmetric. The maximum speed (minimum pressure) along the floor occurs at x* ! 01, which is the same distance as the height of the nozzle above the
V/L
•
r1
V/L 2pr2
Floor →
V
(10)
where V is the magnitude of the resultant velocity vector along the floor as sketched in Fig. 10–72. Since we have made the irrotational flow approximation, the Bernoulli equation can be used to generate the pressure field. Ignoring gravity,
Bernoulli equation:
y
b
•
V/L
r2
x
•
V/L 2pr1
Image source
FIGURE 10–72 Vector sum of the velocities induced by the two sources; the resultant velocity is horizontal at any location on the x-axis due to symmetry.
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510 FLUID MECHANICS 0.2
0.005 Axial velocity
0.15
0
0.1
–0.005
0.05 u*
Pressure coefficient
0
–0.01 Cp –0.015
–0.05
FIGURE 10–73 Nondimensional axial velocity (blue curve) and pressure coefficient (black curve) along the floor below a vacuum cleaner modeled as an irrotational region of flow.
–0.025
–0.15
–0.03
–0.2 –5
•
–4
–3
–2 –1 0 1 2 3 Normalized distance along the floor, x*
4
5
floor (Fig. 10– 74). In dimensional terms, the maximum speed along the floor occurs at x ! &b, and the speed there is
V y Vacuum nozzle
w
Maximum speed along the floor: # V ,L %0.314 m2,s 0u 0 max ! % 0u* 0 max ! %0.159a b ! 2.50 m,s b 0.020 m
b b
–0.02
–0.1
b x Stagnation point Maximum speed
FIGURE 10–74 Based on an irrotational flow approximation, the maximum speed along the floor beneath a vacuum cleaner nozzle occurs at x ! 0b. A stagnation point occurs directly below the nozzle.
(16)
We expect that the vacuum cleaner is most effective at sucking up dirt from the floor when the speed along the floor is greatest and the pressure along the floor is lowest. Thus, contrary to what you may have thought, the best performance is not directly below the suction inlet, but rather at x ! 0b, as illustrated in Fig. 10–74. Discussion Notice that we never used the width w of the vacuum nozzle in our analysis, since a line sink has no length scale. You can convince yourself that a vacuum cleaner works best at x ≅ 0b by performing a simple experiment with a vacuum cleaner and some small granular material (like sugar or salt) on a hard floor. It turns out that the irrotational approximation is quite realistic for flow into the inlet of a vacuum cleaner everywhere except very close to the floor, because the flow is rotational there.
We conclude this section by emphasizing that although the irrotational flow approximation is mathematically simple, and velocity and pressure fields are easy to obtain, we must be very careful where we apply it. The irrotational flow approximation breaks down in regions of non-negligible vorticity, especially near solid walls, where fluid particles rotate because of viscous stresses caused by the no-slip condition at the wall. This leads us to the final section in this chapter (Section 10–6) in which we discuss the boundary layer approximation.
10–6
■
THE BOUNDARY LAYER APPROXIMATION
As discussed in Sections 10–4 and 10–5, there are at least two flow situations in which the viscous terms in the Navier–Stokes equation can be neglected. The first occurs in high Reynolds number regions of flow where net viscous forces are known to be negligible compared to inertial and/or
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511 CHAPTER 10
pressure forces; we call these inviscid regions of flow. The second situation occurs when the vorticity is negligibly small; we call these irrotational or potential regions of flow. In either case, removal of the viscous terms from the Navier–Stokes equation yields the Euler equation (Eq. 10–13 and also Eq. 10–25). While the math is greatly simplified by dropping the viscous terms, there are some serious deficiencies associated with application of the Euler equation to practical engineering flow problems. High on the list of deficiencies is the inability to specify the no-slip condition at solid walls. This leads to unphysical results such as zero viscous shear forces on solid walls and zero aerodynamic drag on bodies immersed in a free stream. We can therefore think of the Euler equation and the Navier–Stokes equation as two mountains separated by a huge chasm (Fig. 10–75a). We make the following statement about the boundary layer approximation: The boundary layer approximation bridges the gap between the Euler equation and the Navier–Stokes equation, and between the slip condition and the no-slip condition at solid walls (Fig. 10–75b).
From a historical perspective, by the mid-1800s, the Navier–Stokes equation was known, but couldn’t be solved except for flows of very simple geometries. Meanwhile, mathematicians were able to obtain beautiful analytical solutions of the Euler equation and of the potential flow equations for flows of complex geometry, but their results were often physically meaningless. Hence, the only reliable way to study fluid flows was empirically, i.e., with experiments. A major breakthrough in fluid mechanics occurred in 1904 when Ludwig Prandtl (1875–1953) introduced the boundary layer approximation. Prandtl’s idea was to divide the flow into two regions: an outer flow region that is inviscid and/or irrotational, and an inner flow region called a boundary layer—a very thin region of flow near a solid wall where viscous forces and rotationality cannot be ignored (Fig. 10–76). In the outer flow region, we use the continuity and Euler equations to obtain the outer flow velocity field, and the Bernoulli equation to obtain the pressure field. Alternatively, if the outer flow region is irrotational, we may use the potential flow techniques discussed in Section 10–5 (e.g., superposition) to obtain the outer flow velocity field. In either case, we solve for the outer flow region first, and then fit in a thin boundary layer in regions where rotationality and viscous forces cannot be neglected. Within the boundary layer we solve the boundary layer equations, to be discussed shortly. (Note that the boundary layer equations are themselves approximations of the full Navier–Stokes equation, as we will see.) The boundary layer approximation corrects some of the major deficiencies of the Euler equation by providing a way to enforce the no-slip condition at solid walls. Hence, viscous shear forces can exist along walls, bodies immersed in a free stream can experience aerodynamic drag, and flow separation in regions of adverse pressure gradient can be predicted more accurately. The boundary layer concept therefore became the workhorse of engineering fluid mechanics throughout most of the 1900s. However, the advent of fast, inexpensive computers and computational fluid dynamics software in the latter part of the twentieth century enabled numerical solution of the Navier–Stokes equation for flows of complex geometry. Today, therefore, it is no longer necessary to split the flow into outer flow regions and boundary layer regions—we can use CFD to solve the full set of equations of motion
No slip
Slip Euler equation
Navier– Stokes equation
(a)
Boundary layer approximation No slip
Slip Euler equation
Navier– Stokes equation
(b)
FIGURE 10–75 (a) A huge gap exists between the Euler equation (which allows slip at walls) and the Navier–Stokes equation (which supports the no-slip condition); (b) the boundary layer approximation bridges that gap.
y
V
Outer flow (inviscid and/or irrotational region of flow) x
d(x)
Boundary layer (rotational with non-negligible viscous forces)
FIGURE 10–76 Prandtl’s boundary layer concept splits the flow into an outer flow region and a thin boundary layer region (not to scale).
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512 FLUID MECHANICS y
Rex ~ 102 d(x) x
V
(a) y
Rex ~ 104 d(x)
x
V (b)
FIGURE 10–77 Flow of a uniform stream parallel to a flat plate (drawings not to scale): (a) Rex ! 102, (b) Rex ! 104. The larger the Reynolds number, the thinner the boundary layer along the plate at a given x-location.
FIGURE 10–78 Flow visualization of a laminar flat plate boundary layer profile. Photograph taken by F. X. Wortmann in 1953 as visualized with the tellurium method. Flow is from left to right, and the leading edge of the flat plate is far to the left of the field of view. Wortmann, F. X. 1977 AGARD Conf. Proc. no. 224, paper 12.
(continuity plus Navier–Stokes) throughout the whole flow field. Nevertheless, boundary layer theory is still useful in some engineering applications, since it takes much less time to arrive at a solution. In addition, there is a lot we can learn about the behavior of flowing fluids by studying boundary layers. We stress again that boundary layer solutions are only approximations of full Navier–Stokes solutions, and we must be careful where we apply this or any approximation. The key to successful application of the boundary layer approximation is the assumption that the boundary layer is very thin. The classic example is a uniform stream flowing parallel to a long flat plate aligned with the x-axis. Boundary layer thickness d at some location x along the plate is sketched in Fig. 10–77. By convention, d is usually defined as the distance away from the wall at which the velocity component parallel to the wall is 99 percent of the fluid speed outside the boundary layer. It turns out that for a given fluid and plate, the higher the free-stream speed V, the thinner the boundary layer (Fig. 10–77). In nondimensional terms, we define the Reynolds number based on distance x along the wall, Reynolds number along a flat plate:
Rex !
rVx Vx ! m n
(10–60)
Hence, At a given x-location, the higher the Reynolds number, the thinner the boundary layer.
In other words, the higher the Reynolds number, the thinner the boundary layer, all else being equal, and the more reliable the boundary layer approximation. We are confident that the boundary layer is thin when d '' x (or, expressed nondimensionally, d/x '' 1). The shape of the boundary layer profile can be obtained experimentally by flow visualization. An example is shown in Fig. 10–78 for a laminar boundary layer on a flat plate. Taken over 50 years ago by F. X. Wortmann, this is now considered a classic photograph of a laminar flat plate boundary
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513 CHAPTER 10
layer profile. The no-slip condition is clearly verified at the wall, and the smooth increase in flow speed away from the wall verifies that the flow is indeed laminar. Note that although we are discussing boundary layers in connection with the thin region near a solid wall, the boundary layer approximation is not limited to wall-bounded flow regions. The same equations may be applied to free shear layers such as jets, wakes, and mixing layers (Fig. 10–79), provided that the Reynolds number is sufficiently high that these regions are thin. The regions of these flow fields with non-negligible viscous forces and finite vorticity can also be considered to be boundary layers, even though a solid wall boundary may not even be present. Boundary layer thickness d(x) is labeled in each of the sketches in Fig. 10–79. As you can see, by convention d is usually defined based on half of the total thickness of the free shear layer. We define d as the distance from the centerline to the edge of the boundary layer where the change in speed is 99 percent of the maximum change in speed from the centerline to the outer flow. Boundary layer thickness is not a constant, but varies with downstream distance x. In the examples discussed here (flat plate, jet, wake, and mixing layer), d(x) increases with x. There are flow situations however, such as rapidly accelerating outer flow along a wall, in which d(x) decreases with x. A common misunderstanding among beginning students of fluid mechanics is that the curve representing d as a function of x is a streamline of the flow—it is not! In Fig. 10–80 we sketch both streamlines and d(x) for the boundary layer growing on a flat plate. As the boundary layer thickness grows downstream, streamlines passing through the boundary layer must diverge slightly upward in order to satisfy conservation of mass. The amount of this upward displacement is smaller than the growth of d(x). Since streamlines cross the curve d(x), d(x) is clearly not a streamline (streamlines cannot cross each other or else mass would not be conserved). For a laminar boundary layer growing on a flat plate, as in Fig. 10–80, boundary layer thickness d is at most a function of V, x, and fluid properties r and m. It is a simple exercise in dimensional analysis to show that d/x is a function of Rex. In fact, it turns out that d is proportional to the square root of Rex. You must note, however, that these results are valid only for a laminar boundary layer on a flat plate. As we move down the plate to larger and larger values of x, Rex increases linearly with x. At some point, infinitesimal disturbances in the flow begin to grow, and the boundary layer cannot remain laminar—it begins a transition process toward turbulent flow. For a smooth flat plate with a uniform free stream, the transition process begins at a critical Reynolds number, Rex, critical ≅ 1 * 105, and continues until the boundary layer is fully turbulent at the transition Reynolds number, Rex, transition ≅ 3 * 106 (Fig. 10–81). The transition process is quite complicated, and details are beyond the scope of this text. V
y
Streamlines
d(x)
d(x) x
Boundary layer
d(x) x
(a) V d(x) x
V (b) V2 d(x) x V1 (c)
FIGURE 10–79 Three additional flow regions where the boundary layer approximation may be appropriate: (a) jets, (b) wakes, and (c) mixing layers.
FIGURE 10–80 Comparison of streamlines and the curve representing d as a function of x for a flat plate boundary layer. Since streamlines cross the curve d(x), d(x) cannot itself be a streamline of the flow.
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514 FLUID MECHANICS
y
V
d(x)
FIGURE 10–81 Transition of the laminar boundary layer on a flat plate into a fully turbulent boundary layer (not to scale).
x Laminar
Transitional
Rex # 105
Turbulent
Rex # 3 ! 106
y
FIGURE 10–82 Thickness of the boundary layer on a flat plate, drawn to scale. Laminar, transitional, and turbulent regions are indicated for the case of a smooth wall with calm free-stream conditions.
y V
Trip wire
x Laminar
Transitional
Turbulent
FIGURE 10–83 A trip wire is often used to initiate early transition to turbulence in a boundary layer (not to scale).
d(x)
Laminar
5
V
Turbulent Transitional
0 0
5
10
15
20
25
30
35
40
x
Note that in Fig. 10–81 the vertical scale has been greatly exaggerated, and the horizontal scale has been shortened (in reality, since Rex, transition ≅ 30 times Rex, critical, the transitional region is much longer than indicated in the figure). To give you a better feel for how thin a boundary layer actually is, we have plotted d as a function of x to scale in Fig. 10–82. To generate the plot, we carefully selected the parameters such that Rex ! 100,000x regardless of the units of x. Thus, Rex, transition occurs at x ≅ 1 and Rex, critical occurs at x ≅ 30 in the plot. Notice how thin the boundary layer is and how long the transitional region is when plotted to scale. In real-life engineering flows, transition to turbulent flow usually occurs more abruptly and much earlier (at a lower value of Rex) than the values given for a smooth flat plate with a calm free stream. Factors such as roughness along the surface, free-stream disturbances, acoustic noise, flow unsteadiness, vibrations, and curvature of the wall contribute to an earlier transition location. Because of this, an engineering critical Reynolds number of Rex, cr ! 5 * 105 is often used to determine whether a boundary layer is most likely laminar (Rex ' Rex, cr ) or most likely turbulent (Rex . Rex, cr ). It is also common in heat transfer to use this value as the critical Re; in fact, relations for average friction and heat transfer coefficients are derived by assuming the flow to be laminar for Rex lower than Rex, cr , and turbulent otherwise. The logic here is to ignore transition by treating the first part of transition as laminar and the remaining part as turbulent. We follow this convention throughout the rest of the book unless noted otherwise. The transition process is often unsteady as well and is difficult to predict, even with modern CFD codes. In some cases, engineers install rough sandpaper or wires called trip wires along the surface, in order to force transition at a desired location (Fig. 10–83). The eddies from the trip wire cause enhanced local mixing and create disturbances that very quickly lead to a turbulent boundary layer. Again, the vertical scale in Fig. 10–83 is greatly exaggerated for illustrative purposes.
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EXAMPLE 10–9
Laminar or Turbulent Boundary Layer?
An aluminum canoe moves horizontally along the surface of a lake at 5.0 mi/h (Fig. 10–84). The temperature of the lake water is 50°F. The bottom of the canoe is 16 ft long and is flat. Is the boundary layer on the canoe bottom laminar or turbulent?
SOLUTION We are to assess whether the boundary layer on the bottom of a canoe is laminar or turbulent. Assumptions 1 The flow is steady and incompressible. 2 Ridges, dings, and other nonuniformities in the bottom of the canoe are ignored; the bottom is assumed to be a smooth flat plate aligned exactly with the direction of flow. 3 From the frame of reference of the canoe, the water below the boundary layer under the canoe moves at uniform speed V ! 5.0 mi/h. Properties The kinematic viscosity of water at T ! 50°F is n ! 1.407 * 10%5 ft2/s. Analysis First, we calculate the Reynolds number at the stern of the canoe using Eq. 10–60, Rex !
(5.0 mi,h) (16 ft) 5280 ft Vx h ! a ba b ! 8.34 * 10 6 n mi 3600 s 1.407 * 10 %5 ft2,s
V
x Boundary layer
d(x)
FIGURE 10–84 Boundary layer growing along the flat bottom of a canoe. Boundary layer thickness is exaggerated for clarity.
(1)
Since Rex is much greater than Rex, cr , and is even greater than Rex, transition, the boundary layer is definitely turbulent by the back of the canoe. Discussion Since the canoe bottom is not perfectly smooth or perfectly flat, and since we expect some disturbances in the lake water due to waves, the paddles, swimming fish, etc., transition to turbulence is expected to occur much earlier and more rapidly than illustrated for the ideal case in Fig. 10–81. Hence, we are even more confident that this boundary layer is turbulent.
The Boundary Layer Equations
Now that we have a physical feel for boundary layers, we need to generate the equations of motion to be used in boundary layer calculations—the boundary layer equations. For simplicity we consider only steady, twodimensional flow in the xy-plane in Cartesian coordinates. The methodology used here can be extended, however, to axisymmetric boundary layers or to three-dimensional boundary layers in any coordinate system. We neglect gravity since we are not dealing with free surfaces or with buoyancy-driven flows (free convection flows), where gravitational effects dominate. We consider only laminar boundary layers; turbulent boundary layer equations are beyond the scope of this text. For the case of a boundary layer along a solid wall, we adopt a coordinate system in which x is everywhere parallel to the wall and y is everywhere normal to the wall (Fig. 10–85). This coordinate system is called a boundary layer coordinate system. When we solve the boundary layer equations, we do so at one x-location at a time, using this coordinate system locally, and it is locally orthogonal. It is not critical where we define x ! 0, but for flow over a body, as in Fig. 10–85, we typically set x ! 0 at the front stagnation point.
V y y
x
Boundary layer y
x x
x=0 L
FIGURE 10–85 The boundary layer coordinate system for flow over a body; x follows the surface and is typically set to zero at the front stagnation point of the body, and y is everywhere normal to the surface locally.
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516 FLUID MECHANICS
We begin with the nondimensionalized Navier–Stokes equation derived at the beginning of this chapter. With the unsteady term and the gravity term neglected, Eq. 10–6 becomes → → → → → 1 (V * $ § *)V * ! %[Eu]§ *P* # c d§*2V * Re
Boundary layer U = U(x)
d y x
Wall
FIGURE 10–86 Magnified view of the boundary layer along the surface of a body, showing length scales x and d and velocity scale U.
The Euler number is of order unity, since pressure differences outside the boundary layer are determined by the Bernoulli equation and 1P ! P % P& ! rV 2. We note that V is a characteristic velocity scale of the outer flow, typically equal to the free-stream velocity for bodies immersed in a uniform flow. The characteristic length scale used in this nondimensionalization is L, some characteristic size of the body. For boundary layers, x is of order of magnitude L, and the Reynolds number in Eq. 10–61 can be thought of as Rex (Eq. 10–60). Rex is very large in typical applications of the boundary layer approximation. It would seem then that we could neglect the last term in Eq. 10–61 in boundary layers. However, doing so would result in the Euler equation, along with all its deficiencies discussed previously. So, we must keep at least some of the viscous terms in Eq. 10–61. How do we decide which terms to keep and which to neglect? To answer this question, we redo the nondimensionalization of the equations of motion based on appropriate length and velocity scales within the boundary layer. A magnified view of a portion of the boundary layer of Fig. 10–85 is sketched in Fig. 10–86. Since the order of magnitude of x is L, we use L as an appropriate length scale for distances in the streamwise direction and for derivatives of velocity and pressure with respect to x. However, this length scale is much too large for derivatives with respect to y. It makes more sense to use d as the length scale for distances in the direction normal to the streamwise direction and for derivatives with respect to y. Similarly, while the characteristic velocity scale is V for the whole flow field, it is more appropriate to use U as the characteristic velocity scale for boundary layers, where U is the magnitude of the velocity component parallel to the wall at a location just above the boundary layer (Fig. 10–86). U is in general a function of x. Thus, within the boundary layer at some value of x, the orders of magnitude are u!U
y
v
"u "v # !0 "x "y
u
d
" 1 ! "x L
P % P& ! rU 2
" 1 ! "y d
(10–62)
The order of magnitude of velocity component v is not specified in Eq. 10–62, but is instead obtained from the continuity equation. Applying the orders of magnitude in Eq. 10–62 to the incompressible continuity equation in two dimensions,
U
Boundary layer
(10–61)
F
FIGURE 10–87 Highly magnified view of the boundary layer along the surface of a body, showing that velocity component v is much smaller than u.
U v ! L d
F
x Wall
→
!U/L
!v/d
Since the two terms have to balance each other, they must be of the same order of magnitude. Thus we obtain the order of magnitude of velocity component v, v!
Ud L
(10–63)
Since d/L '' 1 in a boundary layer (the boundary layer is very thin), we conclude that v '' u in a boundary layer (Fig. 10–87). From Eqs. 10–62
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517 CHAPTER 10
and 10–63, we define the following nondimensional variables within the boundary layer: x* !
x L
y* !
y d
u* !
u U
v* !
vL Ud
P* !
P % P& rU 2
Since we used appropriate scales, all these nondimensional variables are of order unity—i.e., they are normalized variables (Chap. 7). We now consider the x- and y-components of the Navier–Stokes equation. We substitute these nondimensional variables into the y-momentum equation, giving u
" v*Ud "y* Ld
#
n
"2v "x 2
2 1 " P*rU r "y* d
n
"2 v*Ud "x*2
L3
#
n
"2v "y 2
V
Ud L
1 "P % r "y
V
v*
!
V
" v*Ud "x* L2
"v "y
F
F
u*U
# v }
}
"v "x
n
"2 v*Ud "y*2 Ld 2
After some algebra and after multiplying each term by L2/(U2d), we get u*
"v* "v* L 2 "P* n "2v* n L 2 "2v* # a ba b # v* ! %a b #a b (10–64) 2 "x* "y* d "y* UL "x* UL d "y*2
Comparing terms in Eq. 10–64, the middle term on the right side is clearly orders of magnitude smaller than any other term since ReL ! UL/n .. 1. For the same reason, the last term on the right is much smaller than the first term on the right. Neglecting these two terms leaves the two terms on the left and the first term on the right. However, since L .. d, the pressure gradient term is orders of magnitude greater than the advective terms on the left side of the equation. Thus, the only term left in Eq. 10–64 is the pressure term. Since no other term in the equation can balance that term, we have no choice but to set it equal to zero. Thus, the nondimensional ymomentum equation reduces to "P* "0 "y*
Boundary layer
P2 P2
or, in terms of the physical variables, Normal pressure gradient through a boundary layer:
Outer flow
"P "0 "y
(10–65)
In words, although pressure may vary along the wall (in the x-direction), there is negligible change in pressure in the direction normal to the wall. This is illustrated in Fig. 10–88. At x ! x1, P ! P1 at all values of y across the boundary layer from the wall to the outer flow. At some other x-location, x ! x2, the pressure may have changed, but P ! P2 at all values of y across that portion of the boundary layer. The pressure across a boundary layer (y-direction) is nearly constant.
Physically, because the boundary layer is so thin, streamlines within the boundary layer have negligible curvature when observed at the scale of the boundary layer thickness. Curved streamlines require a centripetal acceleration, which comes from a pressure gradient along the radius of curvature. Since the streamlines are not significantly curved in a thin boundary layer, there is no significant pressure gradient across the boundary layer.
P1 P1 P1 P1
y
P2 d
P2
x x2 Wall
x1
FIGURE 10–88 Pressure may change along a boundary layer (x-direction), but the change in pressure across a boundary layer (y-direction) is negligible.
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518 FLUID MECHANICS Boundary layer
Outer flow
P2
y P1
d P1 x1
P2 x2
x Wall
Pressure taps P1
P2
FIGURE 10–89 The pressure in the irrotational region of flow outside of a boundary layer can be measured by static pressure taps in the surface of the wall. Two such pressure taps are sketched.
One immediate consequence of Eq. 10–65 and the statement just presented is that at any x-location along the wall, the pressure at the outer edge of the boundary layer (y ≅ d) is the same as that at the wall (y ! 0). This leads to a tremendous practical application; namely, the pressure at the outer edge of a boundary layer can be measured experimentally by a static pressure tap at the wall directly beneath the boundary layer (Fig. 10–89). Experimentalists routinely take advantage of this fortunate situation, and countless airfoil shapes for airplane wings and turbomachinery blades were tested with such pressure taps over the past century. The experimental pressure data shown in Fig. 10–64 for flow over a circular cylinder were measured with pressure taps at the cylinder’s surface, yet they are used to compare with the pressure calculated by the irrotational outer flow approximation. Such a comparison is valid, because the pressure obtained outside of the boundary layer (from the Euler equation or potential flow analysis coupled with the Bernoulli equation) applies all the way through the boundary layer to the wall. Returning to the development of the boundary layer equations, we use Eq. 10–65 to greatly simplify the x-component of the momentum equation. Specifically, since P is not a function of y, we replace "P/"x by dP/dx, where P is the value of pressure calculated from our outer flow approximation (using either continuity plus Euler, or the potential flow equations plus Bernoulli). The x-component of the Navier–Stokes equation becomes u
#
2 1 " P*rU r "x* L
" u*U "y* d
n
"2u "x 2
n
#
n
"2u "y 2
V
Ud L
1 dP % r dx
V
v*
!
V
" u*U "x* L
"u "y
F
F
u*U
# v }
}
"u "x
"2 u*U "x*2 L2
n
"2 u*U "y*2 d 2
After some algebra, and after multiplying each term by L/U2, we get u*
d(x) ~ √x V
(10–66)
Comparing terms in Eq. 10–66, the middle term on the right side is clearly orders of magnitude smaller than the terms on the left side, since ReL ! UL/n .. 1. What about the last term on the right? If we neglect this term, we throw out all the viscous terms and are back to the Euler equation. Clearly this term must remain. Furthermore, since all the remaining terms in Eq. 10–66 are of order unity, the combination of parameters in parentheses in the last term on the right side of Eq. 10–66 must also be of order unity,
U(x) = V
y
"u* dP* n "2u* L 2 "2u* n "u* # v* !% #a b # a ba b 2 "x* "y* dx* UL "x* UL d "y*2
d(x) x
L 2 n ba b !1 UL d
a
Again recognizing that ReL ! UL/n, we see immediately that 1 d ! L 2ReL
FIGURE 10–90 An order-of-magnitude analysis of the laminar boundary layer equations along a flat plate reveals that d grows like 1x (not to scale).
(10–67)
This confirms our previous statement that at a given streamwise location along the wall, the larger the Reynolds number, the thinner the boundary layer. If we substitute x for L in Eq. 10–67, we also conclude that for a laminar boundary layer on a flat plate, where U(x) ! V ! constant, d grows like the square root of x (Fig. 10–90).
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In terms of the original (physical) variables, Eq. 10–66 is written as x-momentum boundary layer equation: u
"u "u 1 dP "2u #v !% #n r dx "x "y "y 2
→
1 dP dU ! %U r dx dx
(10–69)
where we note that both P and U are functions of x only, as illustrated in Fig. 10–91. Substitution of Eq. 10–69 into Eq. 10–68 yields u
(10–70)
"u "v # !0 "x "y 2
u
y P1
U1
x2
Wall
x
Boundary layer
x1
FIGURE 10–91 Outer flow speed parallel to the wall is U(x) and is obtained from the outer flow pressure, P(x). This speed appears in the x-component of the boundary layer momentum equation, Eq. 10–70. No boundary conditions on downstream edge of flow domain u = U(x)
"u dU "2u "u #v !U #n "x "y dx "y 2
and we have eliminated pressure from the boundary layer equations. We summarize the set of equations of motion for a steady, incompressible, laminar boundary layer in the xy-plane without significant gravitational effects,
Boundary layer equations:
U2
P2
d(x)
(10–68)
Note that the last term in Eq. 10–68 is not negligible in the boundary layer, since the y-derivative of velocity gradient "u/"y is sufficiently large to offset the (typically small) value of kinematic viscosity n. Finally, since we know from our y-momentum equation analysis that the pressure across the boundary layer is the same as that outside the boundary layer (Eq. 10–65), we apply the Bernoulli equation to the outer flow region. Differentiating with respect to x we get P 1 2 # U ! constant r 2
P = P(x), U = U(x)
"u dU "u "u #v !U #n "x "y dx "y 2
(10–71)
Mathematically, the full Navier–Stokes equation is elliptic in space, which means that boundary conditions are required over the entire boundary of the flow domain. Physically, flow information is passed in all directions, both upstream and downstream. On the other hand, the x-momentum boundary layer equation (the second equation of Eq. 10–71) is parabolic. This means that we need to specify boundary conditions on only three sides of the (two-dimensional) flow domain. Physically, flow information is not passed in the direction opposite to the flow (from downstream). This fact greatly reduces the level of difficulty in solving the boundary layer equations. Specifically, we don’t need to specify boundary conditions downstream, only upstream and on the top and bottom of the flow domain (Fig. 10–92). For a typical boundary layer problem along a wall, we specify the no-slip condition at the wall (u ! v ! 0 at y ! 0), the outer flow condition at the edge of the boundary layer and beyond [u ! U(x) as y → &], and a starting profile at some upstream location [u ! ustarting(y) at x ! xstarting, where xstarting may or may not be zero]. With these boundary conditions, we simply march downstream in the x-direction, solving the boundary layer equations as we go. This is particularly attractive for numerical boundary layer computations, because once we know the profile at one x-location (xi), we can march to the next x-location (xi#1), and then use this newly calculated profile as the starting profile to march to the next x-location (xi#2).
y Flow domain x xstarting
u=v=0
u = ustarting(y)
FIGURE 10–92 The boundary layer equation set is parabolic, so boundary conditions need to be specified on only three sides of the flow domain.
Step 1: Calculate U(x) (outer flow).
Step 2: Assume a thin boundary layer.
Step 3: Solve boundary layer equations.
Step 4: Calculate quantities of interest.
Step 5: Verify that boundary layer is thin.
FIGURE 10–93 Summary of the boundary layer procedure for steady, incompressible, two-dimensional boundary layers in the xy-plane.
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The Boundary Layer Procedure
When the boundary layer approximation is employed, we use a general step-by-step procedure. We outline the procedure here and in condensed form in Fig. 10–93.
y
U(x)
d
Boundary layer x
R
Wall
FIGURE 10–94 When the local radius of curvature of the wall (R) is small enough to be of the same magnitude as d, centripetal acceleration effects cannot be ignored and "P/"y - 0. The thin boundary layer approximation is not appropriate in such regions.
Step 1 Solve for the outer flow, ignoring the boundary layer (assuming that the region of flow outside the boundary layer is approximately inviscid and/or irrotational). Transform coordinates as necessary to obtain U(x). Step 2 Assume a thin boundary layer—so thin, in fact, that it does not affect the outer flow solution of step 1. Step 3 Solve the boundary layer equations (Eqs. 10–71), using appropriate boundary conditions: the no-slip boundary condition at the wall, u ! v ! 0 at y ! 0; the known outer flow condition at the edge of the boundary layer, u → U(x) as y → &; and some known starting profile, u ! ustarting(y) at x ! xstarting. Step 4 Calculate quantities of interest in the flow field. For example, once the boundary layer equations have been solved (step 3), we can calculate d(x), shear stress along the wall, total skin friction drag, etc. Step 5 Verify that the boundary layer approximations are appropriate. In other words, verify that the boundary layer is thin—otherwise the approximation is not justified. Before we do any examples, we list here some of the limitations of the boundary layer approximation. These are red flags to look for when performing boundary layer calculations: • The boundary layer approximation breaks down if the Reynolds number is not large enough. How large is large enough? It depends on the desired accuracy of the approximation. Using Eq. 10–67 as a guideline, d/L ! 0.03 (3 percent) for ReL ! 1000, and d/L ! 0.01 (1 percent) for ReL ! 10,000. • The assumption of zero pressure gradient in the y-direction (Eq. 10– 65) breaks down if the wall curvature is of similar magnitude as d (Fig. 10–94). In such cases, centripetal acceleration effects due to streamline curvature cannot be ignored. Physically, the boundary layer is not thin enough for the approximation to be appropriate when d is not '' R. • When the Reynolds number is too high, the boundary layer does not remain laminar, as discussed previously. The boundary layer approximation itself may still be appropriate, but Eqs. 10–71 are not valid if the flow is transitional or fully turbulent. As noted before, the laminar boundary layer on a smooth flat plate under clean flow conditions begins to transition toward turbulence at Rex ≅ 1 * 105. In practical engineering applications, walls may not be smooth and there may be vibrations, noise, and fluctuations in the free-stream flow above the wall, all of which contribute to an even earlier start of the transition process. • If flow separation occurs, the boundary layer approximation is no longer appropriate in the separated flow region. The main reason for this is that a separated flow region contains reverse flow, and the parabolic nature of the boundary layer equations is lost.
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521 CHAPTER 10
EXAMPLE 10–10
V
Laminar Boundary Layer on a Flat Plate
A uniform free stream of speed V flows parallel to an infinitesimally thin semi-infinite flat plate as sketched in Fig. 10–95. The coordinate system is defined such that the plate begins at the origin. Since the flow is symmetric about the x-axis, only the upper half of the flow is considered. Calculate the boundary layer velocity profile along the plate and discuss.
SOLUTION We are to calculate the boundary layer velocity profile (u as a function of x and y) as the laminar boundary layer grows along the flat plate. Assumptions 1 The flow is steady, incompressible, and two-dimensional in the xy-plane. 2 The Reynolds number is high enough that the boundary layer approximation is reasonable. 3 The boundary layer remains laminar over the range of interest. Analysis We follow the step-by-step procedure outlined in Fig. 10–93. Step 1 The outer flow is obtained by ignoring the boundary layer altogether, since it is assumed to be very, very thin. Recall that any streamline in an irrotational flow can be thought of as a wall. In this case, the x-axis can be thought of as a streamline of uniform free-stream flow, one of our building block flows in Section 10–5; this streamline can also be thought of as an infinitesimally thin plate (Fig. 10–96). Thus, U(x) ! V ! constant
Outer flow:
(1)
y
r, m, n Infinitesimally thin flat plate x
FIGURE 10–95 Setup for Example 10–10; flow of a uniform stream parallel to a semiinfinite flat plate along the x-axis.
V
y U(x) = V
x
FIGURE 10–96 The outer flow of Example 10–10 is trivial since the x-axis is a streamline of the flow, and U(x) ! V ! constant.
For convenience, we use U instead of U(x) from here on, since it is a constant. Step 2 We assume a very thin boundary layer along the wall (Fig. 10–97). The key here is that the boundary layer is so thin that it has negligible effect on the outer flow calculated in step 1. Step 3 We must now solve the boundary layer equations. We see from Eq. 1 that dU/dx ! 0; in other words, no pressure gradient term remains in the x-momentum boundary layer equation. This is why the boundary layer on a flat plate is often called a zero pressure gradient boundary layer. The continuity and x-momentum equations for the boundary layer (Eqs. 10–71) become
"u "v # !0 "x "y
u
"u "u "2u #v !n 2 "x "y "y
(2)
There are four required boundary conditions,
u!0
at y ! 0
v ! 0 at y ! 0
u ! U as y → & u ! U for all y at x ! 0
(3)
The last of the boundary conditions in Eq. 3 is the starting profile; we assume that the plate has not yet influenced the flow at the starting location of the plate (x ! 0). These equations and boundary conditions seem simple enough, but unfortunately no analytical solution has ever been found. However, Eqs. 2 were first solved numerically in 1908 by P. R. Heinrich Blasius (1883–1970). As a side note, Blasius was a Ph.D. student of Prandtl. In those days, of course, computers were not yet available, and all the calculations were performed by hand. Today we can solve these equations
V
y U(x) = V
Boundary layer x
FIGURE 10–97 The boundary layer is so thin that it does not affect the outer flow; boundary layer thickness is exaggerated here for clarity.
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522 FLUID MECHANICS U(x) = V
y Magnifying glass or zoom tool
V
d(x) u x
+
(a)
on a computer in a few seconds. The key to the solution is the assumption of similarity. In simple terms, similarity can be assumed here because there is no characteristic length scale in the geometry of the problem. Physically, since the plate is infinitely long in the x-direction, we always see the same flow pattern no matter how much we zoom in or zoom out (Fig. 10–98). Blasius introduced a similarity variable h that combines independent variables x and y into one nondimensional independent variable,
h!y U(x) = V
y
V
d(x)
f( !
x
FIGURE 10–98 A useful result of the similarity assumption is that the flow looks the same (is similar) regardless of how far we zoom in or out; (a) view from a distance, as a person might see, (b) close-up view, as an ant might see.
(4)
and he solved for a nondimensionalized form of the x-component of velocity,
u (b)
U B nx
u ! function of h U
(5)
When we substitute Eqs. 4 and 5 into Eqs. 2, subjected to the boundary conditions of Eq. 3, we get an ordinary differential equation for nondimensional speed f ((h) ! u/U as a function of similarity variable h. We use the popular Runge–Kutta numerical technique to obtain the results shown in Table 10–3 and in Fig. 10–99. Details of the numerical technique are beyond the scope of this text (see Heinsohn and Cimbala, 2003). There is also a small y-component of velocity v away from the wall, but v '' u, and is not discussed here. The beauty of the similarity solution is that this one unique velocity profile shape applies to any x-location when plotted in similarity variables, as in Fig. 10–99. The agreement of the calculated profile shape in Fig. 10–99 to experimentally obtained data (circles in Fig. 10–99) and to the visualized profile shape of Fig. 10–78 is remarkable. The Blasius solution is a stunning success.
TA B L E 1 0 – 3 Solution of the Blasius laminar flat plate boundary layer in similarity variables* h 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 2.2
f((
f(
f
0.33206 0.33205 0.33198 0.33181 0.33147 0.33091 0.33008 0.32739 0.32301 0.31659 0.30787 0.29666 0.28293 0.26675 0.24835
0.00000 0.03321 0.06641 0.09960 0.13276 0.16589 0.19894 0.26471 0.32978 0.39378 0.45626 0.51676 0.57476 0.62977 0.68131
0.00000 0.00166 0.00664 0.01494 0.02656 0.04149 0.05973 0.10611 0.16557 0.23795 0.32298 0.42032 0.52952 0.65002 0.78119
h 2.4 2.6 2.8 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 8.0 9.0 10.0
f(( 0.22809 0.20645 0.18401 0.16136 0.10777 0.06423 0.03398 0.01591 0.00658 0.00240 0.00077 0.00022 0.00001 0.00000 0.00000
f( 0.72898 0.77245 0.81151 0.84604 0.91304 0.95552 0.97951 0.99154 0.99688 0.99897 0.99970 0.99992 1.00000 1.00000 1.00000
f 0.92229 1.07250 1.23098 1.39681 1.83770 2.30574 2.79013 3.28327 3.78057 4.27962 4.77932 5.27923 6.27921 7.27921 8.27921
* h is the similarity variable defined in Eq. 4 above, and function f(h) is solved using the Runge–Kutta numerical technique. Note that f 2 is proportional to the shear stress t, f( is proportional to the x-component of velocity in the boundary layer (f( ! u/U), and f itself is proportional to the stream function. f ( is plotted as a function of h in Fig. 10–99.
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523 CHAPTER 10
Step 4 We next calculate several quantities of interest in this boundary layer. First, based on a numerical solution with finer resolution than that shown in Table 10–3, we find that u/U ! 0.990 at h # 4.91. This 99 percent boundary layer thickness is sketched in Fig. 10–99. Using Eq. 4 and the definition of d, we conclude that y ! d when
U h ! 4.91 ! d B nx
4.91 D ! x 2Rex
→
(6)
This result agrees qualitatively with Eq. 10–67, obtained from a simple order-of-magnitude analysis. The constant 4.91 in Eq. 6 is rounded to 5.0 by many authors, but we prefer to express the result to three significant digits for consistency with other quantities obtained from the Blasius profile. Another quantity of interest is the shear stress at the wall tw,
tw ! m
"u ≤ "y y!0
(7)
Sketched in Fig. 10–99 is the slope of the nondimensional velocity profile at the wall (y ! 0 and h ! 0). From our similarity results (Table 10–3), the nondimensional slope at the wall is
d (u/U) ≤ ! f 2(0) ! 0.332 dh h!0
(8)
6 5 99% boundary layer thickness
4 h
3 2 1
Slope at the wall
0 0
0.2
0.4 0.6 f ' = u/U
0.8
1
FIGURE 10–99 The Blasius profile in similarity variables for the boundary layer growing on a semi-infinite flat plate. Experimental data (circles) are at Rex ! 3.64 * 105. From Panton (1996).
After substitution of Eq. 8 into Eq. 7 and some algebra (transformation of similarity variables back to physical variables), we obtain
tw ! 0.332
Shear stress in physical variables:
rU 2
(9)
2Rex
Thus, we see that the wall shear stress decays with x like x%1/2, as sketched in Fig. 10–100. At x ! 0, Eq. 9 predicts that tw is infinite, which is unphysical. The boundary layer approximation is not appropriate at the leading edge (x ! 0), because the boundary layer thickness is not small compared to x. Furthermore, any real flat plate has finite thickness, and there is a stagnation point at the front of the plate, with the outer flow accelerating quickly to U(x) ! V. We may ignore the region very close to x ! 0 without loss of accuracy in the rest of the flow. Equation 9 is nondimensionalized by defining a skin friction coefficient (also called a local friction coefficient),
Local friction coefficient, laminar flat plate:
Cf, x ! 1
Tw
2 RU
2
!
0.664 2Rex
(10)
Notice that Eq. 10 for Cf, x has the same form as Eq. 6 for d/x, but with a different constant—both decay like the inverse of the square root of Reynolds number. In Chap. 11, we integrate Eq. 10 to obtain the total friction drag on a flat plate of length L.
V
y
Boundary layer
d(x)
U(x) = V
u tw
x
tw
tw (∂u/∂y)y = 0
tw
FIGURE 10–100 For a laminar flat plate boundary layer, wall shear stress decays like x%1/2 as the slope "u/"y at the wall decreases downstream. The front portion of the plate contributes more skin friction drag than does the rear portion.
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524 FLUID MECHANICS V
Step 5 We need to verify that the boundary layer is thin. Consider the practical example of flow over the hood of your car (Fig. 10–101) while you are driving downtown at 20 mi/h on a hot day. The kinematic viscosity of the air is n ! 1.8 " 10#4 ft2/s. We approximate the hood as a flat plate of length 3.5 ft moving horizontally at a speed of V ! 20 mi/h. First, we approximate the Reynolds number at the end of the hood using Eq. 10–60,
Boundary layer y
U(x) x Hood
Rex !
h Vx (20 mi/h) (3.5 ft) 5280 ft ! b a b ! 5.7 " 10 5 a n mi 3600 ft 1.8 " 10 #4 ft2/s
Since Rex is very close to the ballpark critical Reynolds number, Rex, cr ! 5 " 105, the assumption of laminar flow may or may not be appropriate. Nevertheless, we use Eq. 6 to estimate the thickness of the boundary layer, assuming that the flow remains laminar,
FIGURE 10–101 The boundary layer growing on the hood of a car. Boundary layer thickness is exaggerated for clarity.
d!
4.91x 2Rex
!
4.91(3.5 ft) 25.7 " 10
5
a
12 in b ! 0.27 in ft
(11)
By the end of the hood the boundary layer is only about a quarter of an inch thick, and our assumption of a very thin boundary layer is verified. Discussion The Blasius boundary layer solution is valid only for flow over a flat plate perfectly aligned with the flow. However, it is often used as a quick approximation for the boundary layer developing along solid walls that are not necessarily flat nor exactly parallel to the flow, as in the car hood. As illustrated in step 5, it is not difficult in practical engineering problems to achieve Reynolds numbers greater than the critical value for transition to turbulence. You must be careful not to apply the laminar boundary layer solution presented here when the boundary layer becomes turbulent.
Displacement Thickness
V
Outer flow streamline
As was shown in Fig. 10–80, streamlines within and outside a boundary layer must bend slightly outward away from the wall in order to satisfy conservation of mass as the boundary layer thickness grows downstream. This is because the y-component of velocity, v, is small but finite and positive. Outside of the boundary layer, the outer flow is affected by this deflection of the streamlines. We define displacement thickness d* as the distance that a streamline just outside of the boundary layer is deflected, as sketched in Fig. 10–102. U(x) = V
d*(x)
Displacement thickness is the distance that a streamline just outside of the boundary layer is deflected away from the wall due to the effect of the boundary layer.
y d(x)
Boundary layer
FIGURE 10–102 Displacement thickness defined by a streamline outside of the boundary layer. Boundary layer thickness is exaggerated.
x
We generate an expression for d* for the boundary layer along a flat plate by performing a control volume analysis using conservation of mass. The details are left as an exercise for the reader; the result at any x-location along the plate is Displacement thickness:
d* !
!
0
$
u a1 # b dy U
(10–72)
Note that the upper limit of the integral in Eq. 10–72 is shown as $, but since u ! U everywhere above the boundary layer, it is necessary to inte-
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525 CHAPTER 10
grate only out to some finite distance above d. Obviously d* grows with x as the boundary layer grows (Fig. 10–103). For a laminar flat plate, we integrate the numerical (Blasius) solution of Example 10–10 to obtain 1.72 d* ! x 2Rex
Displacement thickness, laminar flat plate:
Displacement thickness is the imaginary increase in thickness of the wall, as seen by the outer flow, due to the effect of the growing boundary layer.
If we were to solve the Euler equation for the flow around this imaginary thicker plate, the outer flow velocity component U(x) would differ from the original calculation. We could then use this apparent U(x) to improve our boundary layer analysis. You can imagine a modification to the boundary layer procedure of Fig. 10–93 in which we go through the first four steps, calculate d*(x), and then go back to step 1, this time using the imaginary (thicker) body shape to calculate an apparent U(x). Following this, we resolve the boundary layer equations. We could repeat the loop as many times as necessary until convergence. In this way, the outer flow and the boundary layer would be more consistent with each other. The usefulness of this interpretation of displacement thickness becomes obvious if we consider uniform flow entering a channel bounded by two parallel walls (Fig. 10–105). As the boundary layers grow on the upper and lower walls, the irrotational core flow must accelerate to satisfy conservation of mass (Fig. 10–105a). From the point of view of the core flow between the boundary layers, the boundary layers cause the channel walls to appear to converge—the apparent distance between the walls decreases as x increases. This imaginary increase in thickness of one of the walls is equal to d*(x), and the apparent U(x) of the core flow must increase accordingly, as sketched, to satisfy conservation of mass. Boundary layer
Boundary layer
Core flow
y
Apparent U(x) d*(x) d(x)
d(x) x (a)
y
d*(x) d(x)
(10–73)
The equation for d* is the same as that for d, but with a different constant. In fact, for laminar flow over a flat plate, d* at any x-location turns out to be approximately three times smaller than d at that same x-location (Fig. 10–103). There is an alternative way to explain the physical meaning of d* that turns out to be more useful for practical engineering applications. Namely, we can think of displacement thickness as an imaginary or apparent increase in thickness of the wall from the point of view of the inviscid and/or irrotational outer flow region. For our flat plate example, the outer flow no longer “sees” an infinitesimally thin flat plate; rather it sees a finite-thickness plate shaped like the displacement thickness of Eq. 10–73, as illustrated in Fig. 10–104.
y
U(x) = V V
x (b)
x
Boundary layer
FIGURE 10–103 For a laminar flat plate boundary layer, the displacement thickness is roughly one-third of the 99 percent boundary layer thickness. Apparent U(x) y
d(x)
V
d*(x)
Boundary layer
Actual wall
x
Apparent wall
FIGURE 10–104 The boundary layer affects the outer flow in such a way that the wall appears to take the shape of the displacement thickness. The apparent U(x) differs from the original approximation.
FIGURE 10–105 The effect of boundary layer growth on flow entering a two-dimensional channel: the irrotational flow between the top and bottom boundary layers accelerates as indicated by (a) actual velocity profiles, and (b) change in apparent core flow due to the displacement thickness of the boundary layer (boundary layers greatly exaggerated for clarity).
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526 FLUID MECHANICS Flow straighteners
EXAMPLE 10–11
Diffuser
Test section
Displacement Thickness in the Design of a Wind Tunnel
Silencer Fan
V
FIGURE 10–106 Schematic diagram of the wind tunnel of Example 10–11.
A small low-speed wind tunnel (Fig. 10–106) is being designed for calibration of hot wires. The air is at 19°C. The test section of the wind tunnel is 30 cm in diameter and 30 cm in length. The flow through the test section must be as uniform as possible. The wind tunnel speed ranges from 1 to 8 m/s, and the design is to be optimized for an air speed of V ! 4.0 m/s through the test section. (a) For the case of nearly uniform flow at 4.0 m/s at the test section inlet, by how much will the centerline air speed accelerate by the end of the test section? (b) Recommend a design that will lead to a more uniform test section flow.
SOLUTION The acceleration of air through the round test section of a wind tunnel is to be calculated, and a redesign of the test section is to be recommended. Assumptions 1 The flow is steady and incompressible. 2 The walls are smooth, and disturbances and vibrations are kept to a minimum. 3 The boundary layer is laminar. Properties The kinematic viscosity of air at 19°C is n ! 1.507 " 10#5 m2/s. Analysis (a) The Reynolds number at the end of the test section is approximately Rex !
R
R – d*
d* (a)
(b)
FIGURE 10–107 Cross-sectional views of the test section of the wind tunnel of Example 10–11: (a) beginning of test section and (b) end of test section.
(4.0 m/s)(0.30 m) Vx ! ! 7.96 " 10 4 n 1.507 " 10 #5 m2/s
Since Rex is lower than the engineering critical Reynolds number, Rex, cr ! 5 " 105, and is even lower than Rex, critical ! 1 " 105, and since the walls are smooth and the flow is clean, we may assume that the boundary layer on the wall remains laminar throughout the length of the test section. As the boundary layer grows along the wall of the wind tunnel test section, air in the region of irrotational flow in the central portion of the test section accelerates as in Fig. 10–105 in order to satisfy conservation of mass. We use Eq. 10–73 to estimate the displacement thickness at the end of the test section,
d* "
1.72x 2Rex
!
1.72(0.30 m) 27.96 " 10 4
! 1.83 " 10 #3 m ! 1.83 mm
(1)
Two cross-sectional views of the test section are sketched in Fig. 10–107, one at the beginning and one at the end of the test section. The effective radius at the end of the test section is reduced by d* as calculated by Eq. 1. We apply conservation of mass to calculate the average air speed at the end of the test section,
Vend A end ! Vbeginning A beginning
→
Vend ! Vbeginning
pR2 p(R # d*)2
(2)
which yields
Vend ! (4.0 m/s)
(0.15 m)2 ! 4.10 m/s (0.15 m # 1.83 " 10 #3 m)2
(3)
Thus the air speed increases by approximately 2.5 percent through the test section, due to the effect of displacement thickness. (b) What recommendation can we make for a better design? One possibility is to design the test section as a slowly diverging duct, rather than as a
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527 CHAPTER 10
straight-walled cylinder (Fig. 10–108). If the radius were designed so as to increase like d*(x) along the length of the test section, the displacement effect of the boundary layer would be eliminated, and the test section air speed would remain fairly constant. Note that there is still a boundary layer growing on the wall, as illustrated in Fig. 10–108. However, the core flow speed outside the boundary layer remains constant, unlike the situation of Fig. 10–105. The diverging wall recommendation would work well at the design operating condition of 4.0 m/s and would help somewhat at other flow speeds. Another option is to apply suction along the wall of the test section in order to remove some of the air along the wall. The advantage of this design is that the suction can be carefully adjusted as wind tunnel speed is varied so as to ensure constant air speed through the test section at any operating condition. This recommendation is the more complicated, and probably more expensive, option. Discussion Wind tunnels have been constructed that use either the diverging wall option or the wall suction option to carefully control the uniformity of the air speed through the wind tunnel test section. The same displacement thickness technique is applied to larger wind tunnels, where the boundary layer is turbulent; however, a different equation for d*(x) is required.
Original test section wall d*(x)
V
d(x)
Core flow
x
Modified test section wall Original test section wall d*(x)
V
d(x)
Apparent core flow
x
Momentum Thickness
Another measure of boundary layer thickness is momentum thickness, commonly given the symbol u. Momentum thickness is best explained by analyzing the control volume of Fig. 10–109 for a flat plate boundary layer. Since the bottom of the control volume is the plate itself, no mass or momentum can cross that surface. The top of the control volume is taken as a streamline of the outer flow. Since no flow can cross a streamline, there can be no mass or momentum flux across the upper surface of the control volume. When we apply conservation of mass to this control volume, we find that the mass flow entering the control volume from the left (at x ! 0) must equal the mass flow exiting from the right (at some arbitrary location x along the plate), 0!
!
→
→
rV & n dA ! wr
CS
!
Y%d*
u dy # wr
at location x
U dy
(10–74) V
0
at x ! 0
where w is the width into the page in Fig. 10–109, which we take arbitrarily as unit width, and Y is the distance from the plate to the outer streamline at x ! 0, as indicated in Fig. 10–109. Since u ! U ! constant everywhere along the left surface of the control volume, and since u ! U between y ! Y and y ! Y % d* along the right surface of the control volume, Eq. 10–74 reduces to
!
Y
(U # u) dy ! Ud*
FIGURE 10–108 A diverging test section would eliminate flow acceleration due to the displacement effect of the boundary layer: (a) actual flow and (b) apparent irrotational core flow.
Y
0
!
Modified test section wall
(10–75)
0
Physically, the mass flow deficit within the boundary layer (the lower blueshaded region in Fig. 10–109) is replaced by a chunk of free-stream flow of thickness d* (the upper blue-shaded region in Fig. 10–109). Equation 10–75 verifies that these two shaded regions have the same area. We zoom in to show these areas more clearly in Fig. 10–110.
y
Outer flow streamline
d*(x) U(x) = V
Y
d(x) FD, x
u
x Boundary layer
FIGURE 10–109 A control volume is defined by the thick dashed line, bounded above by a streamline outside of the boundary layer, and bounded below by the flat plate; FD, x is the viscous force of the plate acting on the control volume.
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528 FLUID MECHANICS y U(x) d*(x)
Free-stream mass flow
Mass flow deficit due to boundary layer
u d(x) x
Now consider the x-component of the control volume momentum equation. Since no momentum crosses the upper or lower control surfaces, the net force acting on the control volume must equal the momentum flux exiting the control volume minus that entering the control volume, Conservation of x-momentum for the control volume: a Fx ! %FD, x !
$
→
→
ruV $ n dA ! rw
CS
u 2 dy % rw
0
at location x
$
Y
U 2 dy
(10–76)
0
x
Wall
$
Y#d*
at x ! 0
where FD, x is the drag force due to friction on the plate from x ! 0 to location x. After some algebra, including substitution of Eq. 10–75, Eq. 10–76 reduces to
FIGURE 10–110 Comparison of the area under the boundary layer profile, representing the mass flow deficit, and the area generated by a chunk of free-stream fluid of thickness d*. To satisfy conservation of mass, these two areas must be identical.
FD, x ! rw
$
Y
u(U % u) dy
(10–77)
0
Finally, we define momentum thickness u such that the viscous drag force on the plate per unit width into the page is equal to rU2 times u, i.e., FD, x w
!r
$
Y
u (U % u) dy % rU 2u
(10–78)
0
In words, Momentum thickness is defined as the loss of momentum flux per unit width divided by rU 2 due to the presence of the growing boundary layer.
Equation 10–78 reduces to u!
$
Y
0
u u a1 % b dy U U
(10–79)
Streamline height Y can be any value, as long as the streamline taken as the upper surface of the control volume is above the boundary layer. Since u ! U for any y greater than Y, we may replace Y by infinity in Eq. 10–79 with no change in the value of u, Momentum thickness:
u!
$
0
&
u u a1 % b dy U U
(10–80)
For the specific case of the Blasius solution for a laminar flat plate boundary layer (Example 10–10), we integrate Eq. 10–80 numerically to obtain U(x) = V
Momentum thickness, laminar flat plate:
V y d(x)
d*(x)
Boundary layer
u(x)
x
FIGURE 10–111 For a laminar flat plate boundary layer, displacement thickness is 35.0 percent of d, and momentum thickness is 13.5 percent of d.
u 0.664 ! x 2Re x
(10–81)
We note that the equation for u is the same as that for d or for d* but with a different constant. In fact, for laminar flow over a flat plate, u turns out to be approximately 13.5 percent of d at any x-location, as indicated in Fig. 10–111. It is no coincidence that u/x (Eq. 10–81) is identical to Cf, x (Eq. 10 of Example 10–10)—both are derived from skin friction drag on the plate.
Turbulent Flat Plate Boundary Layer
It is beyond the scope of this text to derive or attempt to solve the turbulent flow boundary layer equations. Expressions for the boundary layer profile
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529 CHAPTER 10
shape and other properties of the turbulent boundary layer are obtained empirically (or at best semi-empirically), since we cannot solve the boundary layer equations for turbulent flow. Note also that turbulent flows are inherently unsteady, and the instantaneous velocity profile shape varies with time (Fig. 10–112). Thus, all turbulent expressions discussed here represent time-averaged values. One common empirical approximation for the timeaveraged velocity profile of a turbulent flat plate boundary layer is the oneseventh-power law, y 1/7 u "a b U d
for y 3 d,
→
u " 1 for y . d U
y d
(10–82)
Note that in the approximation of Eq. 10–82, d is not the 99 percent boundary layer thickness, but rather the actual edge of the boundary layer, unlike the definition of d for laminar flow. Equation 10–82 is plotted in Fig. 10–113. For comparison, the laminar flat plate boundary layer profile (the numerical solution of Blasius, Fig. 10–99) is also plotted in Fig. 10–113, using y/d for the vertical axis in place of similarity variable h. You can see that if the laminar and turbulent boundary layers were the same thickness, the turbulent one would be much fuller than the laminar one. In other words, the turbulent boundary layer would “hug” the wall more closely, filling the boundary layer with higher-speed flow close to the wall. This is due to the large turbulent eddies that transport high-speed fluid from the outer part of the boundary layer down to the lower parts of the boundary layer (and vice versa). In other words, a turbulent boundary layer has a much greater degree of mixing when compared to a laminar boundary layer. In the laminar case, fluid mixes slowly due to viscous diffusion. However, the large eddies in a turbulent flow promote much more rapid and thorough mixing. The approximate turbulent boundary layer velocity profile shape of Eq. 10–82 is not physically meaningful very close to the wall (y → 0) since it predicts that the slope ("u/"y) is infinite at y ! 0. While the slope at the wall is very large for a turbulent boundary layer, it is nevertheless finite. This large slope at the wall leads to a very high wall shear stress, tw ! m("u/"y)y!0, and, therefore, correspondingly high skin friction along the surface of the plate (as compared to a laminar boundary layer of the same thickness). The skin friction drag produced by both laminar and turbulent boundary layers is discussed in greater detail in Chap. 11. A nondimensionalized plot such as that of Fig. 10–113 is somewhat misleading, since the turbulent boundary layer would actually be much thicker than the corresponding laminar boundary layer at the same Reynolds number. This fact is illustrated in physical variables in Example 10–12. We compare in Table 10–4 expressions for d, d*, u, and Cf, x for laminar and turbulent boundary layers on a smooth flat plate. The turbulent expressions are based on the one-seventh-power law of Eq. 10–82. Note that the expressions in Table 10–4 for the turbulent flat plate boundary layer are valid only for a very smooth surface. Even a small amount of surface roughness can greatly affect properties of the turbulent boundary layer, such as momentum thickness and local skin friction coefficient. The effect of surface roughness on a turbulent flat plate boundary layer is discussed in greater detail in Chap. 11.
0
U
u
FIGURE 10–112 Illustration of the unsteadiness of a turbulent boundary layer; the thin, wavy black lines are instantaneous profiles, and the thick blue line is a long time-averaged profile.
1.2 1 0.8 y — d 0.6 Laminar
0.4 0.2
Turbulent 0 0
0.2
0.4 0.6 u/U
0.8
1
FIGURE 10–113 Comparison of laminar and turbulent flat plate boundary layer profiles, nondimensionalized by boundary layer thickness.
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530 FLUID MECHANICS
TA B L E 1 0 – 4 Summary of expressions for laminar and turbulent boundary layers on a smooth flat plate aligned parallel to a uniform stream* (a)
(b)
Turbulent(†)
Turbulent(‡)
Property
Laminar
Boundary layer thickness
4.91 d ! x 2Rex
0.16 d " x (Rex)1/7
0.38 d " x (Rex)1/5
Displacement thickness
1.72 d* ! x 2Rex
0.020 d* " x (Rex)1/7
0.048 d* " x (Rex)1/5
u 0.664 ! x 2Rex
0.016 u " x (Rex)1/7
0.037 u " x (Rex)1/5
Momentum thickness Local skin friction coefficient
C f, x !
0.664 2Rex
C f, x "
0.027 (Rex)1/7
C f, x "
0.059 (Rex)1/5
* Laminar values are exact and are listed to three significant digits, but turbulent values are listed to only two significant digits due to the large uncertainty affiliated with all turbulent flow fields. † Obtained from one-seventh-power law. ‡ Obtained from one-seventh-power law combined with empirical data for turbulent flow through smooth pipes.
EXAMPLE 10–12 y
V
Comparison of Laminar and Turbulent Boundary Layers
U(x) = V
dturbulent
dlaminar
x
L
FIGURE 10–114 Comparison of laminar and turbulent boundary layers for flow of air over a flat plate for Example 10–12 (boundary layer thickness exaggerated).
Air at 20°C flows at V ! 10.0 m/s over a smooth flat plate of length L ! 1.52 m (Fig. 10–114). (a) Plot and compare the laminar and turbulent boundary layer profiles in physical variables (u as a function of y) at x ! L. (b) Compare the values of local skin friction coefficient for the two cases at x ! L. (c) Plot and compare the growth of the laminar and turbulent boundary layers.
SOLUTION We are to compare laminar versus turbulent boundary layer profiles, local skin friction coefficient, and boundary layer thickness at the end of a flat plate. Assumptions 1 The plate is smooth, and the free stream is calm and uniform. 2 The flow is steady in the mean. 3 The plate is infinitesimally thin and is aligned parallel to the free stream. Properties The kinematic viscosity of air at 20°C is n ! 1.516 * 10%5 m2/s. Analysis (a) First we calculate the Reynolds number at x ! L, Rex !
Vx (10.0 m/s)(1.52 m) ! ! 1.00 * 10 6 n 1.516 * 10 %5 m2/s
This value of Rex is in the transitional region between laminar and turbulent, according to Fig. 10–81. Thus, a comparison between the laminar and turbulent velocity profiles is appropriate. For the laminar case, we multiply the y/d values of Fig. 10–113 by dlaminar, where
d laminar !
4.91x 2Rex
!
4.91(1520 mm) 21.00 * 10 6
! 7.46 mm
(1)
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531 CHAPTER 10
This gives us y-values in units of mm. Similarly, we multiply the u/U values of Fig. 10–113 by U (U ! V ! 10.0 m/s) to obtain u in units of m/s. We plot the laminar boundary layer profile in physical variables in Fig. 10–115. We calculate the turbulent boundary layer thickness at this same x-location using the equation provided in Table 10–4, column (a),
y, mm 20
(2)
[The value of dturbulent based on column (b) of Table 10–4 is somewhat higher, namely 36.4 mm.] Comparing Eqs. 1 and 2, we see that the turbulent boundary layer is about 4.5 times thicker than the laminar boundary layer at a Reynolds number of 1.0 * 106. The turbulent boundary layer velocity profile of Eq. 10–82 is converted to physical variables and plotted in Fig. 10–115 for comparison with the laminar profile. The two most striking features of Fig. 10–115 are (1) the turbulent boundary layer is much thicker than the laminar one, and (2) the slope of u versus y near the wall is much steeper for the turbulent case. (Keep in mind, of course, that very close to the wall the one-seventh-power law does not adequately represent the actual turbulent boundary layer profile.) (b) We use the expressions in Table 10–4 to compare the local skin friction coefficient for the two cases. For the laminar boundary layer,
0.664 2Rex
!
0.664 21.00 * 10 6
30 Turbulent
0.16(1520 mm) 0.16x d turbulent " ! ! 33.8 mm (Rex)1/7 (1.00 * 10 6)1/7
C f, x, laminar !
40
! 6.64 * 10%4
(3)
10 Laminar 0 0
2
4 6 u, m/s
8
10
FIGURE 10–115 Comparison of laminar and turbulent flat plate boundary layer profiles in physical variables at the same x-location. The Reynolds number is Rex ! 1.0 * 106.
and for the turbulent boundary layer, column (a),
C f, x, turbulent "
0.027 0.027 ! ! 3.8 * 10%3 (Rex)1/7 (1.00 * 10 6)1/7
(4)
Comparing Eqs. 3 and 4, the turbulent skin friction value is more than five times larger than the laminar value. If we had used the other expression for turbulent skin friction coefficient, column (b) of Table 10–4, we would have obtained Cf, x, turbulent ! 3.7 * 10%3, very close to the value calculated in Eq. 4. (c) The turbulent calculation assumes that the boundary layer is turbulent from the beginning of the plate. In reality, there is a region of laminar flow, followed by a transition region, and then finally a turbulent region, as illustrated in Fig. 10–81. Nevertheless, it is interesting to compare how dlaminar and dturbulent grow as functions of x for this flow, assuming either all laminar flow or all turbulent flow. Using the expressions in Table 10–4, both of these are plotted in Fig. 10–116 for comparison. 40 Turbulent (b)
30
Turbulent (a)
d, mm 20
Laminar
10 0 0
0.5
1 x, m
1.5
FIGURE 10–116 Comparison of the growth of a laminar boundary layer and a turbulent boundary layer for the flat plate of Example 10–12.
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532 FLUID MECHANICS
Discussion The ordinate in Fig. 10–116 is in mm, while the abscissa is in m for clarity—the boundary layer is incredibly thin, even for the turbulent case. The difference between the turbulent (a) and (b) cases (see Table 10–4) is explained by discrepancies between empirical curve fits and semi-empirical approximations used to obtain the expressions in Table 10–4. This reinforces our decision to report turbulent boundary layer values to at most two significant digits. The real value of d will most likely lie somewhere between the laminar and turbulent values plotted in Fig. 10–116 since the Reynolds number by the end of the plate is within the transitional region.
The one-seventh-power law is not the only turbulent boundary layer approximation used by fluid mechanicians. Another common approximation is the log law, a semi-empirical expression that turns out to be valid not only for flat plate boundary layers but also for fully developed turbulent pipe flow velocity profiles (Chap. 8). In fact, the log law turns out to be applicable for nearly all wall-bounded turbulent boundary layers, not just flow over a flat plate. (This fortunate situation enables us to employ the log law approximation close to solid walls in computational fluid dynamics codes, as discussed in Chap. 15.) The log law is commonly expressed in variables nondimensionalized by a characteristic velocity called the friction velocity u*. (Note that most authors use u* instead of u*. We use a subscript to distinguish u*, a dimensional quantity, from u*, which we use to indicate a nondimensional velocity.) The log law:
u 1 yu * #B ! ln n u* k
(10–83)
where Friction velocity:
u* !
tw Br
(10–84)
and k and B are constants; their usual values are k ! 0.40 to 0.41 and B ! 5.0 to 5.5. Unfortunately, the log law suffers from the fact that it does not work very close to the wall (ln 0 is undefined). It also deviates from experimental values close to the boundary layer edge. Nevertheless, Eq. 10–83 applies across nearly the entire turbulent flat plate boundary layer and is useful because it relates the velocity profile shape to the local value of wall shear stress through Eq. 10–84. A clever expression that is valid all the way to the wall was created by D. B. Spalding in 1961 and is called Spalding’s law of the wall, y
V
d(x)
x
U(x) = V
L
FIGURE 10–117 The turbulent boundary layer generated by flow of air over a flat plate for Example 10–13 (boundary layer thickness exaggerated).
yu * [k(u/u *)]2 [k (u/u *)]3 u ! # e %kB ce k(u/u*) % 1 % k(u/u *) % % d (10–85) n u* 2 6
EXAMPLE 10–13
Comparison of Turbulent Boundary Layer Profile Equations
Air at 20°C flows at V ! 10.0 m/s over a smooth flat plate of length L ! 15.2 m (Fig. 10–117). Plot the turbulent boundary layer profile in physical variables (u as a function of y) at x ! L. Compare the profile generated by the
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533 CHAPTER 10
one-seventh-power law, the log law, and Spalding’s law of the wall, assuming that the boundary layer is fully turbulent from the beginning of the plate.
SOLUTION We are to plot the mean boundary layer profile u(y) at the end of a flat plate using three different approximations. Assumptions 1 The plate is smooth, but there are free-stream fluctuations that tend to cause the boundary layer to transition to turbulence sooner than usual—the boundary layer is turbulent from the beginning of the plate. 2 The flow is steady in the mean. 3 The plate is infinitesimally thin and is aligned parallel to the free stream. Properties The kinematic viscosity of air at 20°C is n ! 1.516 * 10%5 m2/s. Analysis First we calculate the Reynolds number at x ! L, Rex !
Vx (10.0 m/s)(15.2 m) ! ! 1.00 * 10 7 n 1.516 * 10 %5 m2/s
This value of Rex is well above the transitional Reynolds number for a flat plate boundary layer (Fig. 10–81), so the assumption of turbulent flow from the beginning of the plate is reasonable. Using the column (a) values of Table 10–4, we estimate the boundary layer thickness and the local skin friction coefficient at the end of the plate,
d"
0.16x ! 0.240 m (Rex)1/7
C f, x "
0.027 ! 2.70 * 10 %3 (Rex)1/7
(1)
We calculate the friction velocity by using its definition (Eq. 10–84) and the definition of Cf, x (left part of Eq. 10 of Example 10–10),
u* !
C f, x tw 2.70 * 10 %3 !U ! (10.0 m/s) ! 0.367 m/s Br B 2 B 2
(2)
where U ! constant ! V everywhere for a flat plate. It is trivial to generate a plot of the one-seventh-power law (Eq. 10–82), but the log law (Eq. 10–83) is implicit for u as a function of y. Instead, we solve Eq. 10–83 for y as a function of u,
y!
n k(u/u* %B) e u*
(3)
Since we know that u varies from 0 at the wall to U at the boundary layer edge, we are able to plot the log law velocity profile in physical variables using Eq. 3. Finally, Spalding’s law of the wall (Eq. 10–85) is also written in terms of y as a function of u. We plot all three profiles on the same plot for comparison (Fig. 10–118). All three are close, and we cannot distinguish the log law from Spalding’s law on this scale. Instead of a physical variable plot with linear axes as in Fig. 10–118, a semi-log plot of nondimensional variables is often drawn to magnify the nearwall region. The most common notation in the boundary layer literature for the nondimensional variables is y# and u# (inner variables or law of the wall variables), where
Law of the wall variables:
y# !
yu* N
u# !
u u*
(4)
As you can see, y# is a type of Reynolds number, and friction velocity u* is used to nondimensionalize both y and u. Figure 10–118 is redrawn in Fig. 10–119 using law of the wall variables. The differences between the three approximations, especially near the wall, are much clearer when plotted in
250
200
150 y, mm 100
Log law Spalding
50
1/7th power
0 0
2
4 6 u, m/s
8
10
FIGURE 10–118 Comparison of turbulent flat plate boundary layer profile expressions in physical variables at Rex ! 1.0 * 107: one-seventh-power approximation, log law, and Spalding’s law of the wall.
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534 FLUID MECHANICS 30 Experimental data
20
FIGURE 10–119 Comparison of turbulent flat plate boundary layer profile expressions in law of the wall variables at Rex ! 1.0 * 107: one-seventh-power approximation, log law, and Spalding’s law of the wall. Typical experimental data and the viscous sublayer equation (u# ! y#) are also shown for comparison.
u+ = y+ u+
1/7th power 10 Spalding Log law 0 1
10
102 y+
103
104
this fashion. Typical experimental data are also plotted in Fig. 10–119 for comparison. Spalding’s formula does the best job overall and is the only expression that follows experimental data near the wall. In the outer part of the boundary layer, the experimental values of u# level off beyond some value of y#, as does the one-seventh-power law. However, both the log law and Spalding’s formula continue indefinitely as a straight line on this semilog plot. Discussion Also plotted in Fig. 10–119 is the linear equation u# ! y#. The region very close to the wall (0 ' y# ' 5 or 6) is called the viscous sublayer. In this region, turbulent fluctuations are suppressed due to the close proximity of the wall, and the velocity profile is nearly linear. Other names for this region are linear sublayer and laminar sublayer. We see that Spalding’s equation captures the viscous sublayer and blends smoothly into the log law. Neither the one-seventh-power law nor the log law are valid this close to the wall.
Boundary layer
(a)
Boundary Layers with Pressure Gradients Boundary layer
(b)
FIGURE 10–120 Boundary layers with nonzero pressure gradients occur in both external flows and internal flows: (a) boundary layer developing along the fuselage of an airplane and into the wake, and (b) boundary layer growing on the wall of a diffuser (boundary layer thickness exaggerated in both cases).
So far we have spent most of our discussion on flat plate boundary layers. Of more practical concern for engineers are boundary layers on walls of arbitrary shape. These include external flows over bodies immersed in a free stream (Fig. 10–120a), as well as some internal flows like the walls of wind tunnels and other large ducts in which boundary layers develop along the walls (Fig. 10–120b). Just as with the zero pressure gradient flat plate boundary layer discussed earlier, boundary layers with nonzero pressure gradients may be laminar or turbulent. We often use the flat plate boundary layer results as ballpark estimates for such things as location of transition to turbulence, boundary layer thickness, skin friction, etc. However, when more accuracy is needed we must solve the boundary layer equations (Eqs. 10–71 for the steady, laminar, two-dimensional case) using the procedure outlined in Fig. 10–93. The analysis is much harder than that for a flat plate since the pressure gradient term (U dU/dx) in the x-momentum equation is nonzero. Such an analysis can quickly get quite involved, especially for the case of three-dimensional flows. Therefore, we discuss only some qualita-
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535 CHAPTER 10
tive features of boundary layers with pressure gradients, leaving detailed solutions of the boundary layer equations to higher-level fluid mechanics textbooks (e.g., Panton, 1996, and White, 1991). First some terminology. When the flow in the inviscid and/or irrotational outer flow region (outside of the boundary layer) accelerates, U(x) increases and P(x) decreases. We refer to this as a favorable pressure gradient. It is favorable or desirable because the boundary layer in such an accelerating flow is usually thin, hugs closely to the wall, and therefore is not likely to separate from the wall. When the outer flow decelerates, U(x) decreases, P(x) increases, and we have an unfavorable or adverse pressure gradient. As its name implies, this condition is not desirable because the boundary layer is usually thicker, does not hug closely to the wall, and is much more likely to separate from the wall. In a typical external flow, such as flow over an airplane wing (Fig. 10–121), the boundary layer in the front portion of the body is subjected to a favorable pressure gradient, while that in the rear portion is subjected to an adverse pressure gradient. If the adverse pressure gradient is strong enough (dP/dx ! %U dU/dx is large), the boundary layer is likely to separate off the wall. Examples of flow separation are shown in Fig. 10–122 for both external and internal flows. In Fig. 10–122a is sketched an airfoil at a moderate angle of attack. The boundary layer remains attached over the entire lower surface of the airfoil, but it separates somewhere near the rear of the upper surface as sketched. The closed streamline indicates a region of recirculating flow called a separation bubble. As pointed out previously, the boundary layer equations are parabolic, meaning that no information can be passed from the downstream boundary. However, separation leads to reverse flow near the wall, destroying the parabolic nature of the flow field, and rendering the boundary layer equations inapplicable.
Adverse
Favorable
FIGURE 10–121 The boundary layer along a body immersed in a free stream is typically exposed to a favorable pressure gradient in the front portion of the body and an adverse pressure gradient in the rear portion of the body.
The boundary layer equations are not valid downstream of a separation point because of reverse flow in the separation bubble.
In such cases, the full Navier–Stokes equations must be used in place of the boundary layer approximation. From the point of view of the boundary layer procedure of Fig. 10–93, the procedure breaks down because the outer flow calculated in step 1 is no longer valid when separation occurs, especially beyond the separation point (compare Fig. 10–121 to Fig. 10–122a). Figure 10–122b shows the classic case of an airfoil at too high of an angle of attack, in which the separation point moves near the front of the airfoil; the separation bubble covers nearly the entire upper surface of the airfoil—a condition known as stall. Stall is accompanied by a loss of lift and a marked
Separation point
(a)
Separation point
(b)
Separation point
(c)
FIGURE 10–122 Examples of boundary layer separation in regions of adverse pressure gradient: (a) an airplane wing at a moderate angle of attack, (b) the same wing at a high angle of attack (a stalled wing), and (c) a wide-angle diffuser in which the boundary layer cannot remain attached and separates on one side.
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536 FLUID MECHANICS y y
U(x)
U(x)
u
d(x)
d(x)
u
x
tw (a)
x
tw (b) y
y
U(x) U(x)
d(x)
u x
tw (c)
At the wall:
u
d(x)
tw = 0
x
(d) y U(x)
d(x) u
Reverse flow
x
tw
increase in aerodynamic drag, as discussed in more detail in Chap. 11. Flow separation may also occur in internal flows, such as in the adverse pressure gradient region of a diffuser (Fig. 10–122c). As sketched, separation often occurs asymmetrically on one side of the diffuser only. As with an airfoil with flow separation, the outer flow calculation in the diffuser is no longer meaningful, and the boundary layer equations are not valid. Flow separation in a diffuser leads to a significant decrease of pressure recovery, and such conditions in a diffuser are also referred to as stall conditions. We can learn a lot about the velocity profile shape under various pressure gradient conditions by examining the boundary layer momentum equation right at the wall. Since the velocity is zero at the wall (no-slip condition), the entire left side of Eq. 10–71 disappears, leaving only the pressure gradient term and the viscous term, which must balance,
(e)
FIGURE 10–123 Comparison of boundary layer profile shape as a function of pressure gradient (dP/dx ! %U dU/dx): (a) favorable, (b) zero, (c) mild adverse, (d) critical adverse (separation point), and (e) large adverse; inflection points are indicated by blue circles, and wall shear stress tw ! m ("u/"y)y!0 is sketched for each case.
"2u dU 1 dP n a 2b ! %U ! dx r dx "y y!0
(10–86)
Under favorable pressure gradient conditions (accelerating outer flow), dU/dx is positive, and by Eq. 10–86, the second derivative of u at the wall is negative, i.e., ("2u/"y2)y!0 ' 0. We know that "2u/"y2 must remain negative as u approaches U(x) at the edge of the boundary layer. Thus, we expect the velocity profile across the boundary layer to be rounded, without any inflection point, as sketched in Fig. 10–123a. Under zero pressure gradient conditions, ("2u/"y2)y!0 is zero, implying a linear growth of u with respect to y near the wall, as sketched in Fig. 10–123b. (This is verified by the Blasius boundary layer profile for the zero pressure gradient boundary layer on a flat plate, as shown in Fig. 10–99.) For adverse pressure gradients, dU/dx is negative and Eq. 10–86 demands that ("2u/"y2)y!0 be positive. However, since "2u/"y2 must be negative as u approaches U(x) at the edge of the boundary layer, there has to be an inflection point ("2u/"y2 ! 0) somewhere in the boundary layer, as illustrated in Fig. 10–123c. The first derivative of u with respect to y at the wall is directly proportional to tw, the wall shear stress [tw ! m ("u/"y)y!0]. Comparison of ("u/"y)y!0 in Fig. 10–123a through c reveals that tw is largest for favorable pressure gradients and smallest for adverse pressure gradients. Boundary layer thickness increases as the pressure gradient changes sign, as also illustrated in Fig. 10–123. If the adverse pressure gradient is large enough, ("u/"y)y!0 can become zero (Fig. 10–123d); this location along a wall is the separation point, beyond which there is reverse flow and a separation bubble (Fig. 10–123e). Notice that beyond the separation point tw is negative due to the negative value of ("u/"y)y!0. As mentioned previously, the boundary layer equations break down in regions of reverse flow. Thus, the boundary layer approximation may be appropriate up to the separation point, but not beyond. We use computational fluid dynamics (CFD) to illustrate flow separation for the case of flow over a bump along a wall. The flow is steady and twodimensional and Fig. 10–124a shows outer flow streamlines generated by a solution of the Euler equation. Without the viscous terms there is no separation, and the streamlines are symmetric fore and aft. As indicated on the figure, the front portion of the bump experiences an accelerating flow and hence a favorable pressure gradient. The rear portion experiences a decelerating flow and an adverse pressure gradient. When the full (laminar)
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537 CHAPTER 10
Flow direction
Adverse Favorable Bump surface
(a) Outer flow
Separation bubble
Bump surface
Approximate location of separation point
Reverse flow
(b)
Dividing streamline
Approximate location of separation point Reverse flow (c)
Approximate location of separation point
Reverse flow (d)
FIGURE 10–124 CFD calculations of flow over a bump: (a) solution of the Euler equation with outer flow streamlines plotted (no flow separation), (b) laminar flow solution showing flow separation on the downstream side of the bump, (c) close-up view of streamlines near the separation point, and (d) close-up view of velocity vectors, same view as (c).
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538 FLUID MECHANICS
Navier–Stokes equation is solved, the viscous terms lead to flow separation off the rear end of the bump, as seen in Fig. 10–124b. Keep in mind that this is a full Navier–Stokes solution, not a boundary layer solution; nevertheless it illustrates the process of flow separation in the boundary layer. The approximate location of the separation point is indicated in Fig. 10–124b, and the dashed black line is a type of dividing streamline. Fluid below this streamline is caught in the separation bubble, while fluid above this streamline continues downstream. A close-up view of streamlines is shown in Fig. 10–124c, and velocity vectors are plotted in Fig. 10–124d using the same close-up view. Reverse flow in the lower portion of the separation bubble is clearly visible. Also, there is a strong y-component of velocity beyond the separation point, and the outer flow is no longer nearly parallel to the wall. In fact, the separated outer flow is nothing like the original outer flow of Fig. 10–124a. This is typical and represents a serious deficiency in the boundary layer approach. Namely, the boundary layer equations may be able to predict the location of the separation point fairly well, but cannot predict anything beyond the separation point. In some cases the outer flow changes significantly upstream of the separation point as well, and the boundary layer approximation gives erroneous results. The boundary layer approximation is only as good as the outer flow solution; if the outer flow is significantly altered by flow separation, the boundary layer approximation will be erroneous.
The boundary layers sketched in Fig. 10–123 and the flow separation velocity vectors plotted in Fig. 10–124 are for laminar flow. Turbulent boundary layers have qualitatively similar behavior, although as discussed previously, the mean velocity profile of a turbulent boundary layer is much fuller than a laminar boundary layer under similar conditions. Thus a stronger adverse pressure gradient is required to separate a turbulent boundary layer. We make the following general statement: Turbulent boundary layers are more resistant to flow separation than are laminar boundary layers exposed to the same adverse pressure gradient.
Experimental evidence for this statement is shown in Fig. 10–125, in which the outer flow is attempting a sharp turn through a 20° angle. The laminar (a)
FIGURE 10–125 Flow visualization comparison of laminar and turbulent boundary layers in an adverse pressure gradient; flow is from left to right. (a) The laminar boundary layer separates at the corner, but (b) the turbulent one does not. Photographs taken by M. R. Head in 1982 as visualized with titanium tetrachloride. Head, M. R. 1982 in Flow Visualization II, W. Merzkirch, ed., 399–403. Washington: Hemisphere.
(b)
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539 CHAPTER 10 Outer flow
Bump surface
FIGURE 10–126 CFD calculation of turbulent flow over the same bump as that of Fig. 10–124. Compared to the laminar result of Fig. 10–124b, the turbulent boundary layer is more resistant to flow separation and does not separate in the adverse pressure gradient region in the rear portion of the bump.
boundary layer (Fig. 10–125a) cannot negotiate the sharp turn, and separates at the corner. The turbulent boundary layer on the other hand (Fig. 10–125b) manages to remain attached around the sharp corner. As another example, flow over the same bump as that of Fig. 10–124 is recalculated, but with turbulence modeled in the simulation. Streamlines generated by the turbulent CFD calculation are shown in Fig. 10–126. Notice that the turbulent boundary layer remains attached (no flow separation), in contrast to the laminar boundary layer that separates off the rear portion of the bump. In the turbulent case, the outer flow Euler solution (Fig. 10–124a) remains valid over the entire surface since there is no flow separation and since the boundary layer remains very thin. A similar situation occurs for flow over bluff objects like spheres. A smooth golf ball, for example, would maintain a laminar boundary layer on its surface, and the boundary layer would separate fairly easily, leading to large aerodynamic drag. Golf balls have dimples (a type of surface roughness) in order to create an early transition to a turbulent boundary layer. Flow still separates from the golf ball surface, but much farther downstream in the boundary layer, resulting in significantly reduced aerodynamic drag. This is discussed in more detail in Chap. 11.
The Momentum Integral Technique for Boundary Layers
In many practical engineering applications, we do not need to know all the details inside the boundary layer; rather we seek reasonable estimates of gross features of the boundary layer such as boundary layer thickness and skin friction coefficient. The momentum integral technique utilizes a control volume approach to obtain such quantitative approximations of boundary layer properties along surfaces with zero or nonzero pressure gradients. The momentum integral technique is straightforward, and in some applications does not require use of a computer. It is valid for both laminar and turbulent boundary layers. We begin with the control volume sketched in Fig. 10–127. The bottom of the control volume is the wall at y ! 0, and the top is at y ! Y, high enough to enclose the entire height of the boundary layer. The control volume is an infinitesimally thin slice of width dx in the x-direction. In accordance with
U(x)
CV Pleft face y
Pright face Y
BL
d(x) x
d(x + dx)
u dx
tw
x + dx
FIGURE 10–127 Control volume (thick dashed black line) used in derivation of the momentum integral equation.
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540 FLUID MECHANICS
the boundary layer approximation, "P/"y ! 0, so we assume that pressure P acts along the entire left face of the control volume, Pleft face ! P
In the general case with nonzero pressure gradient, the pressure on the right face of the control volume differs from that on the left face. Using a firstorder truncated Taylor series approximation (Chap. 9), we set dP dx dx
Pright face ! P #
In a similar manner we write the incoming mass flow rate through the left face as # mleft face ! rw
# mright face ! rw c
m⋅ right face
BL
dx
(10–87)
and the outgoing mass through the right face as
y
x
u dy
0
m⋅ top
m⋅ left face
$
Y
x + dx
FIGURE 10–128 Mass flow balance on the control volume of Fig. 10–127.
$
Y
d a dx
u dy #
0
$
Y
u dyb dxd
(10–88)
0
where w is the width of the control volume into the page in Fig. 10–127. If you prefer, you can set w to unit width; it will cancel out later anyway. Since Eq. 10–88 differs from Eq. 10–87, and since no flow can cross the bottom of the control volume (the wall), mass must flow into or out of the top face of the control volume. We illustrate this in Fig. 10–128 for the case . . . of a growing boundary layer in which m right face ' m left face, and m top is positive (mass flows out). Conservation of mass over the control volume yields d # mtop ! %rw a dx
$
Y
u dyb dx
(10–89)
0
We now apply conservation of x-momentum for the chosen control volume. The x-momentum is brought in through the left face and is removed through the right and top faces of the control volume. The net momentum flux out of the control volume must be balanced by the force due to the shear stress acting on the control volume by the wall and the net pressure force on the control surface, as shown in Fig. 10–127. The steady control volume x-momentum equation is thus a Fx, body #
ignore gravity
a Fx, surface
YwP % YwaP #
dP dxb % w dx tw dx
!
$
→
→
ruV $ n dA #
left face
→
→
ruV $ n dA #
right face
2
u dy
rwc
0
$
0
Y
→
→
ruV $ n dA
top
d u dy # a dx 2
$
$
Y
$
Y
%rw
$
2
u dyb dxd
. m topU
0
where the momentum flux through the top surface of the control volume is taken as the mass flow rate through that surface times U. Some of the terms cancel, and we rewrite the equation as %Y
dP d % tw ! r a dx dx
$
0
Y
u 2 dyb % rU
d a dx
$
0
Y
u dyb
(10–90)
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541 CHAPTER 10
. where we have used Eq. 10–89 for m top, and w and dx cancel from each remaining term. For convenience we note that Y !
$
Y
dy. From the outer
Product rule: d aU dx
0
flow (Euler equation), dP/dx ! %rU dU/dx. After dividing each term in Eq. 10–90 by density r, we get dU U dx
$
Y
0
tw d a dy % ! r dx
$
Y
d u dyb % U a dx 2
0
$
Y
u dyb
(10–91)
0
We simplify Eq. 10–91 by utilizing the product rule of differentiation in reverse (Fig. 10–129). After some rearrangement, Eq. 10–91 becomes d a dx
$
0
Y
dU u(U % u) dyb # dx
$
Y
0
tw (U % u) dy ! r
where we are able to put U inside the integrals since at any given x-location, U is constant with respect to y (U is a function of x only). We multiply and divide the first term by U2 and the second term by U to get d aU 2 dx
$
0
&
u dU u a1 % b dyb # U U U dx
$
&
0
tw u a1 % b dy ! r U
(10–92)
where we have also substituted infinity in place of Y in the upper limit of each integral since u ! U for all y greater than Y, and thus the value of the integral does not change by this substitution. We previously defined displacement thickness d* (Eq. 10–72) and momentum thickness u (Eq. 10–80) for a flat plate boundary layer. In the general case with nonzero pressure gradient, we define d* and u in the same way, except we use the local value of outer flow velocity, U ! U(x), at a given x-location in place of the constant U since U now varies with x. Equation 10–92 can thus be written in more compact form as Kármán integral equation:
tw dU d (U 2 u) # U d* ! r dx dx
(10–93)
Equation 10–93 is called the Kármán integral equation in honor of Theodor von Kármán (1881–1963), a student of Prandtl, who was the first to derive the equation in 1921. An alternate form of Eq. 10–93 is obtained by performing the product rule on the first term, dividing by U2, and rearranging, Kármán integral equation, alternative form:
C f, x 2
!
du u dU # (2 # H) dx U dx
(10–94)
where we define shape factor H as Shape factor:
H!
d* u
(10–95)
and local skin friction coefficient Cf, x as Local skin friction coefficient:
C f, x ! 1
tw
2 rU
2
U
(10–96)
Note that both H and Cf, x are functions of x for the general case of a boundary layer with a nonzero pressure gradient developing along a surface.
Y
$0 u dyb
=
Y
$
d dU Y a u dyb + u dy dx 0 dx 0
$
Product rule in reverse: U
Y d a u dyb = dx 0
$
d aU dx
Y
$0 u dyb
–
dU Y u dy dx 0
$
FIGURE 10–129 The product rule is utilized in reverse in the derivation of the momentum integral equation.
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542 FLUID MECHANICS
We emphasize again that the derivation of the Kármán integral equation and Eqs. 10–93 through 10–96 are valid for any steady incompressible boundary layer along a wall, regardless of whether the boundary layer is laminar, turbulent, or somewhere in between. For the special case of the boundary layer on a flat plate, U(x) ! U ! constant, and Eq. 10–94 reduces to Kármán integral equation, flat plat boundary layer:
EXAMPLE 10–14
C f, x ! 2
du dx
(10–97)
Flat Plate Boundary Layer Analysis Using the Kármán Integral Equation
Suppose we know only two things about the turbulent boundary layer over a flat plate, namely, the local skin friction coefficient (Fig. 10–130), y
U(x) = V V
C f, x "
d(x) u x
d
Cf, x
FIGURE 10–130 The turbulent boundary layer generated by flow over a flat plate for Example 10–14 (boundary layer thickness exaggerated).
0.027 (Rex)1/7
(1)
and the one-seventh-power law approximation for the boundary layer profile shape,
y 1/7 u "a b U d
u " 1 for y . d U
for y 3 d
(2)
Using the definitions of displacement thickness and momentum thickness and employing the Kármán integral equation, estimate how d, d*, and u vary with x.
SOLUTION We are to estimate d, d*, and u based on Eqs 1 and 2.
Assumptions 1 The flow is turbulent, but steady in the mean. 2 The plate is thin and is aligned parallel to the free stream, so that U(x) ! V ! constant. Analysis First we substitute Eq. 2 into Eq. 10–80 and integrate to find momentum thickness,
u!
$
0
&
u u a1 % b dy ! U U
d
y
$ adb 0
1/7
y 1/7 7 a1 % a b b dy ! d d 72
(3)
Similarly, we find displacement thickness by integrating Eq. 10–72,
d* !
$
0
&
u a1 % b dy ! U
$
d
0
y 1/7 1 a1 % a b b dy ! d d 8
(4)
The Kármán integral equation reduces to Eq. 10–97 for a flat plate boundary layer. We substitute Eq. 3 into Eq. 10–97 and rearrange to get
C f, x ! 2
du 14 dd ! dx 72 dx
from which
dd 72 72 ! C ! 0.027(Rex) %1/7 dx 14 f, x 14
(5)
where we have substituted Eq. 1 for the local skin friction coefficient. Equation 5 can be integrated directly, yielding
Boundary layer thickness:
0.16 D " x (Rex)1/7
(6)
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543 CHAPTER 10
Finally, substitution of Eqs. 3 and 4 into Eq. 6 gives approximations for d* and u,
Displacement thickness:
0.020 D* " x (Rex)1/7
(7)
U 0.016 " x (Rex)1/7
(8)
and
Momentum thickness:
CAUTION INTEGRATION REQUIRED
Discussion The results agree with the expressions given in column (a) of Table 10–4 to two significant digits. Indeed, many of the expressions in Table 10–4 were generated with the help of the Kármán integral equation.
While fairly simple to use, the momentum integral technique suffers from a serious deficiency. Namely, we must know (or guess) the boundary layer profile shape in order to apply the Kármán integral equation (Fig. 10–131). For the case of boundary layers with pressure gradients, boundary layer shape changes with x (as illustrated in Fig. 10–123), further complicating the analysis. Fortunately, the shape of the velocity profile does not need to be known precisely, since integration is very forgiving. Several techniques have been developed that utilize the Kármán integral equation to predict gross features of the boundary layer. Some of these techniques, such as Thwaite’s method, do a very good job for laminar boundary layers. Unfortunately, the techniques that have been proposed for turbulent boundary layers have not been as successful. Many of the techniques require the assistance of a computer and are beyond the scope of the present textbook.
FIGURE 10–131 Integration of a known (or assumed) velocity profile is required when using the Kármán integral equation. Contraction Test section
EXAMPLE 10–15
Drag on the Wall of a Wind Tunnel Test Section
V
A boundary layer develops along the walls of a rectangular wind tunnel. The air is at 20°C and atmospheric pressure. The boundary layer starts upstream of the contraction and grows into the test section (Fig. 10–132). By the time it reaches the test section, the boundary layer is fully turbulent. The boundary layer profile and its thickness are measured at both the beginning (x ! x1) and the end (x ! x2) of the bottom wall of the wind tunnel test section. The test section is 1.8 m long and 0.50 m wide (into the page in Fig. 10–132). The following measurements are made:
d 1 ! 4.2 cm
d 2 ! 7.7 cm
(1)
At both locations the boundary layer profile fits better to a one-eighth-power law approximation than to the standard one-seventh-power law approximation,
y 1/8 u "a b U d
for y 3 d
u " 1 for y . d U
Diffuser
x1
x2 Boundary layer (a)
y
U(x) = V d(x) BL
d1
u x1
d2
FD
x
x2 (b)
(2)
Estimate the total skin friction drag force FD acting on the bottom wall of the wind tunnel test section.
SOLUTION We are to estimate the skin friction drag force on the bottom wall of the test section of the wind tunnel (between x ! x1 and x ! x2).
FIGURE 10–132 Boundary layer developing along the wind tunnel walls of Example 10–15: (a) overall view, and (b) close-up view of the bottom wall of the test section (boundary layer thickness exaggerated).
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544 FLUID MECHANICS
Properties For air at 20°C, n ! 1.516 " 10#5 m2/s and r ! 1.204 kg/m3. Assumptions 1 The flow is steady in the mean. 2 The wind tunnel walls diverge slightly to ensure that U(x) ! V ! constant. Analysis First we substitute Eq. 2 into Eq. 10–80 and integrate to find momentum thickness u,
u!
!
0
%
u u a1 # b dy ! U U
!
d
0
y 1/8 y 1/8 4 a b c1 # a b d dy ! d d d 45
(3)
The Kármán integral equation reduces to Eq. 10–97 for a flat plate boundary layer. In terms of the shear stress along the wall, Eq. 10–97 is
1 du tw ! rU 2C f, x ! rU 2 2 dx
(4)
We integrate Eq. 4 from x ! x1 to x ! x2 to find the skin friction drag force,
FD ! w
!
x2
tw dx ! wrU 2
x1
!
x2
x1
du dx ! wrU 2(u 2 # u 1) dx
(5)
where w is the width of the wall into the page in Fig. 10–132. After substitution of Eq. 3 into Eq. 5 we obtain
FD ! wrU 2
4 (d # d 1) 45 2
(6)
Finally, substitution of the given numerical values into Eq. 6 yields the drag force,
FD ! (0.50 m)(1.204 kg/m3)(10.0 m/s)2
4 s2 $ N (0.077 # 0.042) m a b ! 0.19 N 45 kg $ m
Discussion This is a very small force since the newton is itself a small unit of force. The Kármán integral equation would be much more difficult to apply if the outer flow velocity U(x) were not constant.
y
Fluid properties Flat plate r, m
V L
x
FIGURE 10–133 Flow over an infinitesimally thin flat plate of length L. CFD calculations are reported for ReL ranging from 10#1 to 105.
We end this chapter with some illuminating results from CFD calculations of flow over a two-dimensional, infinitesimally thin flat plate aligned with the free stream (Fig. 10–133). In all cases the plate is 1 m long (L ! 1 m), and the fluid is air with constant properties r ! 1.23 kg/m3 and m ! 1.79 " 10#5 kg/m · s. We vary free-stream velocity V so that the Reynolds number at the end of the plate (ReL ! rVL/m) ranges from 10#1 (creeping flow) to 105 (laminar but ready to start transitioning to turbulent). All cases are incompressible, steady, laminar Navier–Stokes solutions generated by a commercial CFD code. In Fig. 10–134, we plot velocity vectors for four Reynolds number cases at three x-locations: x ! 0 (beginning of the plate), x ! 0.5 m (middle of the plate), and x ! 1 m (end of the plate). We also plot streamlines in the vicinity of the plate for each case. In Fig. 10–134a, ReL ! 0.1, and the creeping flow approximation is reasonable. The flow field is nearly symmetric fore and aft—typical of creeping flow over symmetric bodies. Notice how the flow diverges around the plate as if it were of finite thickness. This is due to the large displacement effect caused by viscosity and the no-slip condition. In essence, the flow velocity near the plate is so small that the rest of the flow “sees” it as a blockage around which the flow must be diverted. The y-component of
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545 CHAPTER 10
y Plate L
x
(a) ReL = 1 * 10–1
(b) ReL = 1 * 101
(c) ReL = 1 * 103
(d) ReL = 1 * 105
FIGURE 10–134 CFD calculations of steady, incompressible, two-dimensional laminar flow from left to right over a 1-m-long flat plate of infinitesimal thickness; velocity vectors are shown in the left column at three locations along the plate, and streamlines near the plate are shown in the right column. ReL ! (a) 0.1, (b) 10, (c) 1000, and (d) 100,000; only the upper half of the flow field is solved— the lower half is a mirror image. The computational domain extends hundreds of plate lengths beyond what is shown here in order to approximate “infinite” far-field conditions at the edges of the computational domain.
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546 FLUID MECHANICS d(L) = V
4.91(1 m)
= 0.155 m 1000 U(x) = V L
FIGURE 10–135 Calculation of boundary layer thickness for a laminar boundary layer on a flat plate at ReL ! 1000. This result is compared to the CFDgenerated velocity profile at x ! L shown in Fig. 10–134c at this same Reynolds number.
x
velocity is significant near both the front and rear of the plate. Finally, the influence of the plate extends tens of plate lengths in all directions into the rest of the flow, which is also typical of creeping flows. The Reynolds number is increased by two orders of magnitude to ReL ! 10 in the results shown in Fig. 10–134b. This Reynolds number is too high to be considered creeping flow, but too low for the boundary layer approximation to be appropriate. We notice some of the same features as those of the lower Reynolds number case, such as a large displacement of the streamlines and a significant y-component of velocity near the front and rear of the plate. The displacement effect is not as strong, however, and the flow is no longer symmetric fore and aft. We are seeing the effects of inertia as fluid leaves the end of the flat plate; inertia sweeps fluid into the developing wake behind the plate. The influence of the plate on the rest of the flow is still large, but much less so than for the flow at ReL ! 0.1. In Fig. 10–134c are shown results of the CFD calculations at ReL ! 1000, another increase of two orders of magnitude. At this Reynolds number, inertial effects are starting to dominate over viscous effects throughout the majority of the flow field, and we can start calling this a boundary layer (albeit a fairly thick one). In Fig. 10–135 we calculate the boundary layer thickness using the laminar expression given in Table 10–4. The predicted value of d(L) is about 15 percent of the plate length at ReL ! 1000, which is in reasonable agreement with the velocity vector plot at x ! L in Fig. 10–134c. Compared to the lower Reynolds number cases of Fig. 10–134a and b, the displacement effect is greatly reduced and any trace of fore–aft symmetry is gone. Finally, the Reynolds number is once again increased by two orders of magnitude to ReL ! 100,000 in the results shown in Fig. 10–134d. There is no question about the appropriateness of the boundary layer approximation at this large Reynolds number. The CFD results show an extremely thin boundary layer with negligible effect on the outer flow. The streamlines of Fig. 10–134d are nearly parallel everywhere, and you must look closely to see the thin wake region behind the plate. The streamlines in the wake are slightly farther apart there than in the rest of the flow, because in the wake region, the velocity is significantly less than the free-stream velocity. The ycomponent of velocity is negligible, as is expected in a very thin boundary layer, since the displacement thickness is so small. Profiles of the x-component of velocity are plotted in Fig. 10–136 for each of the four Reynolds numbers of Fig. 10–134, plus some additional cases at other values of ReL. We use a log scale for the vertical axis (y in units of m), since y spans several orders of magnitude. We nondimensionalize the abscissa as u/U so that the velocity profile shapes can be compared. All the profiles have a somewhat similar shape when plotted in this fashion. However, we notice that some of the profiles have a significant velocity overshoot (u . U) near the outer portion of the velocity profile. This is a direct result of the displacement effect and the effect of inertia as discussed before. At very low values of ReL (ReL 3 100), where the displacement effect is most prominent, the velocity overshoot is almost nonexistent. This can be explained by the lack of inertia at these low Reynolds numbers. Without inertia, there is no mechanism to accelerate the flow around the plate; rather, viscosity retards the flow everywhere in the vicinity of the
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547 CHAPTER 10 1000 100 ReL = 10 –1 10
10 0
y, m 1
101 10 2 10 3
0.1 10 4
10 5
0.01
10 6 0.001 0
0.2
0.4
0.6 u/U
0.8
1
1.2
FIGURE 10–136 CFD calculations of steady, incompressible, two-dimensional laminar flow over a flat plate of infinitesimal thickness: nondimensional x velocity component u/U is plotted against vertical distance from the plate, y. Prominent velocity overshoot is observed at moderate Reynolds numbers, but disappears at very low and very high values of ReL.
plate, and the influence of the plate extends tens of plate lengths beyond the plate in all directions. For example, at ReL ! 10%1, u does not reach 99 percent of U until y # 320 m—more than 300 plate lengths above the plate! At moderate values of the Reynolds number (ReL between about 101 and 104), the displacement effect is significant, and inertial terms are no longer negligible. Hence, fluid is able to accelerate around the plate and the velocity overshoot is significant. For example, the maximum velocity overshoot is about 5 percent at ReL ! 102. At very high values of the Reynolds number (ReL 4 105), inertial terms dominate viscous terms, and the boundary layer is so thin that the displacement effect is almost negligible. The small displacement effect leads to very small velocity overshoot. For example, at ReL ! 106 the maximum velocity overshoot is only about 0.4 percent. Beyond ReL ! 106, laminar flow is no longer physically realistic, and the CFD calculations would need to include the effects of turbulence.
SUMMARY The Navier–Stokes equation is difficult to solve, and therefore approximations are often used for practical engineering analyses. As with any approximation, however, we must be sure that the approximation is appropriate in the region of flow being analyzed. In this chapter we examine several approximations and show examples of flow situations in which they are useful. First we nondimensionalize the Navier–Stokes equation, yielding several nondimensional parameters: the Strouhal number (St), Froude number (Fr), Euler number (Eu), and Reynolds number (Re). Furthermore,
for flows without free-surface effects, the hydrostatic pressure component due to gravity can be incorporated into a modified pressure P(, effectively eliminating the gravity term (and the Froude number) from the Navier–Stokes equation. The nondimensionalized Navier–Stokes equation with modified pressure is →
→ → → → → 1 "V * # (V * $ § *)V * ! %[Eu]§ *P(* # c d§*2 V * [St] "t* Re
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548 FLUID MECHANICS
When the nondimensional variables (indicated by *) are of order of magnitude unity, the relative importance of each term in the equation depends on the relative magnitude of the nondimensional parameters. For regions of flow in which the Reynolds number is very small, the last term in the equation dominates the terms on the left side, and hence pressure forces must balance viscous forces. If we ignore inertial forces completely, we make the creeping flow approximation, and the Navier–Stokes equation reduces to →
mation and a full Navier–Stokes solution. We assume that an inviscid and/or irrotational outer flow exists everywhere except in very thin regions close to solid walls or within wakes, jets, and mixing layers. The boundary layer approximation is appropriate for high Reynolds number flows. However, we recognize that no matter how large the Reynolds number, viscous terms in the Navier–Stokes equations are still important within the thin boundary layer, where the flow is rotational and viscous. The boundary layer equations for steady, incompressible, two-dimensional, laminar flow are
→
§ P( " m§ 2V
Creeping flow is foreign to our everyday observations since our bodies, our automobiles, etc., move about at relatively high Reynolds numbers. The lack of inertia in the creeping flow approximation leads to some very interesting peculiarities, as discussed in this chapter. We define inviscid regions of flow as regions where the viscous terms are negligible compared to the inertial terms (opposite of creeping flow). In such regions of flow the Navier–Stokes equation reduces to the Euler equation, →
→ → → → "V ra # (V $ § )V b ! %§ P( "t
In inviscid regions of flow, the Euler equation can be manipulated to derive the Bernoulli equation, valid along streamlines of the flow. Regions of flow in which individual fluid particles do not rotate are called irrotational regions of flow. In such regions, the vorticity of fluid particles is negligibly small, and the viscous terms in the Navier–Stokes equation can be neglected, leaving us again with the Euler equation. In addition, the Bernoulli equation becomes less restrictive, since the Bernoulli constant is the same everywhere, not just along streamlines. A nice feature of irrotational flow is that elementary flow solutions (building block flows) can be added together to generate more complicated flow solutions, a process known as superposition. Since the Euler equation cannot support the no-slip boundary condition at solid walls, the boundary layer approximation is useful as a bridge between an Euler equation approxi-
"u "v # !0 "x "y
and
u
"u "u dU "2u #v !U #n 2 "x "y dx "y
We define several measures of boundary layer thickness, including the 99 percent thickness d, the displacement thickness d*, and the momentum thickness u. These quantities can be calculated exactly for a laminar boundary layer growing along a flat plate, under conditions of zero pressure gradient. As the Reynolds number increases down the plate, the boundary layer transitions to turbulence; semi-empirical expressions are given in this chapter for a turbulent flat plate boundary layer. The Kármán integral equation is valid for both laminar and turbulent boundary layers exposed to arbitrary nonzero pressure gradients, tw dU d d* ! (U 2u) # U r dx dx This equation is useful for “back of the envelope” estimations of gross boundary layer properties such as boundary layer thickness and skin friction. The approximations presented in this chapter are applied to many practical problems in engineering. Potential flow analysis is useful for calculation of airfoil lift (Chap. 11). We utilize the inviscid approximation in the analysis of compressible flow (Chap. 12), open-channel flow (Chap. 13), and turbomachinery (Chap. 14). In cases where these approximations are not justified, or where more precise calculations are required, the Navier–Stokes equations are solved numerically using CFD (Chap. 15).
REFERENCES AND SUGGESTED READING 1. G. T. Yates. “How Microorganisms Move through Water,” American Scientist, 74, pp. 358–365, July–August, 1986.
4. M. Van Dyke. An Album of Fluid Motion. Stanford, CA: The Parabolic Press, 1982.
2. R. J. Heinsohn and J. M. Cimbala. Indoor Air Quality Engineering. New York: Marcel-Dekker, 2003.
5. F. M. White. Viscous Fluid Flow, 2nd ed. New York: McGraw-Hill, 1991.
3. P. K. Kundu. Fluid Mechanics. San Diego, CA: Academic Press, 1990.
6. R. L. Panton. Incompressible Flow, 2nd ed. New York: Wiley, 1996.
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549 CHAPTER 10
APPLICATION SPOTLIGHT
■
Droplet Formation
Guest Authors: James A. Liburdy and Brian Daniels, Oregon State University Droplet formation is a complex interaction of inertial, surface tension, and viscous forces. The actual break-off of a drop from a stream of liquid, although studied for almost 200 years, has still not been fully explained. Droplet-on Demand (DoD) is used for such diverse applications as ink-jet printing and DNA analysis in microscale “lab-on-a-chip” devices. DoD requires very uniform droplet sizes, controlled velocities and trajectories, and a high rate of sequential droplet formation. For example, in ink-jet printing, the typical size of a droplet is 25 to 50 microns (barely visible with the naked eye), the velocities are on the order of 10 m/s, and the droplet formation rate can be higher than 20,000 per second. The most common method for forming droplets involves accelerating a stream of liquid, and then allowing surface tension to induce an instability in the stream, which breaks up into individual droplets. In 1879, Lord Rayleigh developed a classical theory for the instability associated with this break-up; his theory is still widely used today to define droplet break-up conditions. A small perturbation to the surface of the liquid stream sets up an undulating pattern along the length of the stream, which causes the stream to break up into droplets whose size is determined by the radius of the stream and the surface tension of the liquid. However, most DoD systems rely on acceleration of the stream with time-dependent forcing functions in the form of a pressure wave exerted at the inlet of a nozzle. If the pressure wave is very rapid, viscous effects at the walls are negligible, and the potential flow approximation can be used to predict the flow. Two important nondimensional parameters in DoD are the Ohnesorge number Oh ! m/(rssa)1/2 and the Weber number We ! rVa/ss, where a is the radius of the nozzle, ss is the surface tension, and V is the velocity. The Ohnesorge number determines when viscous forces are important relative to surface tension forces. In addition, the nondimensional pressure required to form an unstable fluid stream, Pc ! Pa/ss, is called the capillary pressure, and the associated capillary time scale for droplets to form is tc ! (ra/ss)1/2. When Oh is small, the potential flow approximation is applicable, and the surface shape is controlled by a balance between surface tension and fluid acceleration. Example surfaces of flow emerging from a nozzle are shown in Fig. 10–137a and b. Surface shape depends on the pressure amplitude and the time scale of the perturbation, and is predicted well using the potential flow approximation. When the pressure is large enough and the pulse is fast enough, the surface ripples, and the center forms a jet stream that eventually breaks off into a droplet (Fig. 10–137c). An area of active research is how to control the size and velocity of these droplets, while producing thousands per second. References Rayleigh, Lord, “On the Instability of Jets,” Proc. London Math. Soc., 10, pp. 4–13, 1879. Daniels, B. J., and Liburdy, J. A., “Oscillating Free-Surface Displacement in an Orifice Leading to Droplet Formation,” J. Fluids Engr., 10, pp. 7–8, 2004.
(a)
(b)
(c)
FIGURE 10–137 Droplet formation starts when a surface becomes unstable to a pressure pulse. Shown here are water surfaces in (a) an 800-micron orifice disturbed by a 5000-Hz pulse and (b) a 1200micron orifice disturbed by an 8100Hz pulse. Reflection from the surface causes the image to appear as if the surface wave is both up and down. The wave is axisymmetric, at least for small-amplitude pressure pulses. The higher the frequency, the shorter the wavelength and the smaller the central node. The size of the central node defines the diameter of the liquid jet, which then breaks up into a droplet. (c) Droplet formation from a highfrequency pressure pulse ejected from a 50-micron-diameter orifice. The center liquid stream produces the droplet and is only about 25 percent of the orifice diameter. Ideally, a single droplet forms, but unwanted, “satellite” droplets are often generated along with the main droplet. Courtesy James A. Liburdy and Brian Daniels, Oregon State University. Used by permission.
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550 FLUID MECHANICS
PROBLEMS* General and Introductory Problems, Modified Pressure, Fluid Statics 10–1C Explain the difference between an “exact” solution of the Navier–Stokes equation (as discussed in Chap. 9) and an approximate solution (as discussed in this chapter). 10–2C A box fan sits on the floor of a very large room (Fig. P10–2C). Label regions of the flow field that may be approximated as static. Label regions in which the irrotational approximation is likely to be appropriate. Label regions where the boundary layer approximation may be appropriate. Finally, label regions in which the full Navier– Stokes equation most likely needs to be solved (i.e., regions where no approximation is appropriate). Box fan
FIGURE P10–2C 10–3C Discuss how nondimensionalization of the Navier– Stokes equation is helpful in obtaining approximate solutions. Give an example.
10–7C What is the most important criterion for use of the modified pressure P( rather than the thermodynamic pressure P in a solution of the Navier–Stokes equation? 10–8C Which nondimensional parameter in the nondimensionalized Navier–Stokes equation is eliminated by use of modified pressure instead of actual pressure? Explain. 10–9 Consider flow of water through a small hole in the bottom of a large cylindrical tank (Fig. P10–9). The flow is laminar everywhere. Jet diameter d is much smaller than tank diameter D, but D is of the same order of magnitude as tank height H. Carrie reasons that she can use the fluid statics approximation everywhere in the tank except near the hole, but wants to validate this approximation mathematically. She lets the characteristic velocity scale in the tank be V ! Vtank. The characteristic length scale is tank height H, the characteristic time is the time required to drain the tank tdrain, and the reference pressure difference is rgH (pressure difference from the water surface to the bottom of the tank, assuming fluid statics). Substitute all these scales into the nondimensionalized incompressible Navier–Stokes equation (Eq. 10–6) and verify by order-of-magnitude analysis that for d '' D, only the pressure and gravity terms remain. In particular, compare the order of magnitude of each term and each of the four nondimensional parameters St, Eu, Fr, and Re. (Hint: Vjet ! 1gH .) Under what criteria is Carrie’s approximation appropriate?
10–4C What is the most significant danger associated with an approximate solution of the Navier–Stokes equation? Give an example that is different than the ones given in this chapter. 10–5C What criteria can you use to determine whether an approximation of the Navier–Stokes equation is appropriate or not? Explain. 10–6C In the nondimensionalized incompressible Navier– Stokes equation (Eq. 10–6), there are four nondimensional parameters. Name each one, explain its physical significance (e.g., the ratio of pressure forces to viscous forces), and discuss what it means physically when the parameter is very small or very large.
* Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
D
→
r, m
H
Vtank
g
d Vjet
FIGURE P10–9 10–10 Consider steady, incompressible, laminar, fully developed, planar Poiseuille flow between two parallel, hori-
u
→
g
P x x1
FIGURE P10–10
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zontal plates (velocity and pressure profiles are shown in Fig. P10–10). At some horizontal location x ! x1, the pressure varies linearly with vertical distance z, as sketched. Choose an appropriate datum plane (z ! 0), sketch the profile of modified pressure all along the vertical slice, and shade in the region representing the hydrostatic pressure component. Discuss. 10–11 Consider the planar Poiseuille flow of Prob. 10–10. Discuss how modified pressure varies with downstream distance x. In other words, does modified pressure increase, stay the same, or decrease with x? If P( increases or decreases with x, how does it do so (e.g., linearly, quadratically, exponentially)? Use a sketch to illustrate your answer. 10–12 In Chap. 9 (Example 9–15), we generated an “exact” solution of the Navier–Stokes equation for fully developed Couette flow between two horizontal flat plates (Fig. P10–12), with gravity acting in the negative z-direction (into the page of Fig. P10–12). We used the actual pressure in that example. Repeat the solution for the x-component of velocity u and pressure P, but use the modified pressure in your equations. The pressure is P0 at z ! 0. Show that you get the same result as previously. Discuss. Answers: u ! Vy/h, P ! P0 % rgz V Moving plate Fluid: r, m
h
Fixed plate
y, v x, u
FIGURE P10–12 10–13 Write out the three components of the Navier–Stokes equation in Cartesian coordinates in terms of modified pressure. Insert the definition of modified pressure and show that the x-, y-, and z-components are identical to those in terms of regular pressure. What is the advantage of using modified pressure?
cated on Fig. P10–14. Sketch the profile of actual pressure all along the vertical slice. Discuss. 10–15 In Example 9–18 we solved the Navier–Stokes equation for steady, fully developed, laminar flow in a round pipe (Poiseuille flow), neglecting gravity. Now, add back the effect of gravity by re-solving that same problem, but use modified pressure P( instead of actual pressure P. Specifically, calculate the actual pressure field and the velocity field. Assume the pipe is horizontal, and let the datum plane z ! 0 be at some arbitrary distance under the pipe. Is the actual pressure at the top of the pipe greater than, equal to, or less than that at the bottom of the pipe? Discuss.
Creeping Flow 10–16C Write a one-word description of each of the five terms in the incompressible Navier–Stokes equation, →
r
→ → → → → "V → # r(V $ § )V ! %§ P # rg # m § 2V "t I
P at a point
→
z x
FIGURE P10–14
P'
III
IV
V
When the creeping flow approximation is made, only two of the five terms remain. Which two terms remain, and why is this significant? 10–17 The viscosity of clover honey is listed as a function of temperature in Table P10–17. The specific gravity of the honey is about 1.42 and is not a strong function of temperature. The honey is squeezed through a small hole of diameter D ! 4.0 mm in the lid of an inverted honey jar. The room and the honey are at T ! 20°C. Estimate the maximum speed of the honey through the hole such that the flow can be approximated as creeping flow. (Assume that Re must be less than 0.1 for the creeping flow approximation to be appropriate.) Repeat your calculation if the temperature is 40°C. Discuss. Answers: 0.33 m/s, 0.035 m/s
TA B L E P 1 0 – 1 7 Viscosity of clover honey at 16 percent moisture content
10–14 A flow field is simulated by a computational fluid dynamics code that uses the modified pressure in its calculations. A profile of modified pressure along a vertical slice through the flow is sketched in Fig. P10–14. The actual pressure at a point midway through the slice is known, as indi-
g
II
T, °C
m, poise*
14 20 30 40 50 70
600 190 65 20 10 3
* Poise ! g/cm · s. Data from Airborne Honey, Ltd., www.airborne.co.nz.
10–18 For each case, calculate an appropriate Reynolds number and indicate whether the flow can be approximated by the creeping flow equations. (a) A microorganism of diameter 5.0 mm swims in room temperature water at a speed
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of 0.2 mm/s. (b) Engine oil at 140°C flows in the small gap of a lubricated automobile bearing. The gap is 0.0012 mm thick, and the characteristic velocity is 20.0 m/s. (c) A fog droplet of diameter 10 mm falls through 30°C air at a speed of 3.0 mm/s.
L
h0
10–19 Estimate the speed and Reynolds number of the sperm shown in Fig. 10–10. Is this microorganism swimming under creeping flow conditions? Assume it is swimming in room-temperature water. 10–20 A good swimmer can swim 100 m in about a minute. If a swimmer’s body is 1.8 m long, how many body lengths does he swim per second? Repeat the calculation for the sperm of Fig. 10–10. In other words, how many body lengths does the sperm swim per second? Use the sperm’s whole body length, not just that of his head, for the calculation. Compare the two results and discuss. 10–21 A drop of water in a rain cloud has diameter D ! 30 mm (Fig. P10–21). The air temperature is 25°C, and its pressure is standard atmospheric pressure. How fast does the air have to move vertically so that the drop will remain suspended in the air? Answer: 0.0264 m/s
D
r, m
V
FIGURE P10–21 10–22 Discuss why fluid density does not influence the aerodynamic drag on a particle moving in the creeping flow regime. 10–23 A slipper-pad bearing (Fig. P10–23) is often encountered in lubrication problems. Oil flows between two blocks; the upper one is stationary, and the lower one is moving in this case. The drawing is not to scale; in actuality, h '' L. The thin gap between the blocks converges with increasing x. Specifically, gap height h decreases linearly from h0 at x ! 0 to hL at x ! L. Typically, the gap height length scale h0 is much smaller than the axial length scale L. This problem is more complicated than simple Couette flow between parallel plates because of the changing gap height. In particular, axial velocity component u is a function of both x and y, and pressure P varies nonlinearly from P ! P0 at x ! 0 to P ! PL at x ! L. ("P/"x is not constant). Gravity forces are negligible in this flow field, which we approximate as two-dimensional, steady, and laminar. In fact, since h is so small and oil is so viscous, the creeping flow approximations
y
u(x, y) x
h(x)
m
hL
V
FIGURE P10–23
are used in the analysis of such lubrication problems. Let the characteristic length scale associated with x be L, and let that associated with y be h0 (x ! L and y ! h0). Let u ! V. Assuming creeping flow, generate a characteristic scale for pressure difference 1P ! P % P0 in terms of L, h0, m, and V. Answer:
mVL /h02
10–24 Consider the slipper-pad bearing of Prob. 10–23. (a) Generate a characteristic scale for v, the y-component of velocity. (b) Perform an order-of-magnitude analysis to compare the inertial terms to the pressure and viscous terms in the x-momentum equation. Show that when the gap is small (h0 '' L) and the Reynolds number is small (Re ! rVh0/m '' 1), the creeping flow approximation is appropriate. (c) Show that when h0 '' L, the creeping flow equations may still be appropriate even if the Reynolds number (Re ! rVh0/m) is not less than 1. Explain. Answer: (a) Vh0/L 10–25 Consider again the slipper-pad bearing of Prob. 10–23. Perform an order-of-magnitude analysis on the ymomentum equation, and write the final form of the ymomentum equation. (Hint: You will need the results of Probs. 10–23 and 10–24.) What can you say about pressure gradient "P/"y? 10–26 Consider again the slipper-pad bearing of Prob. 10–23. (a) List appropriate boundary conditions on u. (b) Solve the creeping flow approximation of the x-momentum equation to obtain an expression for u as a function of y (and indirectly as a function of x through h and dP/dx, which are functions of x). You may assume that P is not a function of y. Your final expression should be written as u(x, y) ! f (y, h, dP/dx, V, and m). Name the two distinct components of the velocity profile in your result. (c) Nondimensionalize your expression for u using these appropriate scales: x* ! x/L, y* ! y/h0, h* ! h/h0, u* ! u/V, and P* ! (P % P0)h02/mVL. 10–27 Consider the slipper-pad bearing of Fig. P10–27. The drawing is not to scale; in actuality, h '' L. This case differs from that of Prob. 10–23 in that h(x) is not linear; rather h is some known, arbitrary function of x. Write an expression for axial velocity component u as a function of y, h, dP/dx, V, and m. Discuss any differences between this result and that of Prob. 10–26.
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10–32 Discuss what happens when the oil temperature increases significantly as the slipper-pad bearing of Prob. 10–31E is subjected to constant use at the amusement park. In particular, would the load-carrying capacity increase or decrease? Why?
L
h0
y
u(x, y) x
h(x)
m
hL
V
FIGURE P10–27 10–28 For the slipper-pad bearing of Prob. 10–23, use the continuity equation, appropriate boundary conditions, and the one-dimensional Leibnitz theorem (see Chap. 4) to show that d dx
$
h
u dy ! 0.
0
10–29 Combine the results of Probs. 10–26 and 10–28 to show that for a two-dimensional slipper-pad bearing, pressure d dP ah 3 b gradient dP/dx is related to gap height h by dx dx
dh . This is the steady, two-dimensional form of the dx more general Reynolds equation for lubrication (Panton, 1996). ! 6mU
10–30 Consider flow through a two-dimensional slipper-pad bearing with linearly decreasing gap height from h0 to hL (Fig. P10–23), namely, h ! h0 # ax, where a is the nondimensional convergence of the gap, a ! (hL % h0)/L. We note that tan a ≅ a for very small values of a. Thus, a is approximately the angle of convergence of the upper plate in Fig. P10–23 (a is negative for this case). Assume that the oil is exposed to atmospheric pressure at both ends of the slipper-pad, so that P ! P0 ! Patm at x ! 0 and P ! PL ! Patm at x ! L. Integrate the Reynolds equation (Prob. 10–29) for this slipper-pad bearing to generate an expression for P as a function of x. 10–31E
A slipper-pad bearing with linearly decreasing gap height (Fig. P10–23) is being designed for an amusement park ride. Its dimensions are h0 ! 1/1000 in (2.54 * 10%5 m), hL ! 1/2000 in (1.27 * 10%5 m), and L ! 1.0 in (0.0254 m). The lower plate moves at speed V ! 10.0 ft/s (3.048 m/s) relative to the upper plate. The oil is engine oil at 40°C. Both ends of the slipper-pad are exposed to atmospheric pressure, as in Prob. 10–30. (a) Calculate the convergence a, and verify that tan a ≅ a for this case. (b) Calculate the gage pressure halfway along the slipper-pad (at x ! 0.5 in). Comment on the magnitude of the gage pressure. (c) Plot P* as a function of x*, where x* ! x/L and P* ! (P % Patm)h02/mVL. (d) Approximately how many pounds of weight (load) can this slipper-pad bearing support if it is b ! 6.0 in deep (into the page of Fig. P10–23)?
10–33 Is the slipper-pad flow of Prob. 10–31E in the creeping flow regime? Discuss. Are the results reasonable? 10–34
We saw in Prob. 10–31E that a slipper-pad bearing can support a large load. If the load were to increase, the gap height would decrease, thereby increasing the pressure in the gap. In this sense, the slipper-pad bearing is “self-adjusting” to varying loads. If the load increases by a factor of 2, calculate how much the gap height decreases. Specifically, calculate the new value of h0 and the percentage change. Assume that the slope of the upper plate and all other parameters and dimensions stay the same as those in Prob. 10–31E. 10–35 Estimate the speed at which you would need to swim in room temperature water to be in the creeping flow regime. (An order-of-magnitude estimate will suffice.) Discuss.
Inviscid Flow 10–36C In what way is the Euler equation an approximation of the Navier–Stokes equation? Where in a flow field is the Euler equation an appropriate approximation? 10–37C What is the main difference between the steady, incompressible Bernoulli equation for irrotational regions of flow, and the steady incompressible Bernoulli equation for rotational but inviscid regions of flow? 10–38 In the derivation of the Bernoulli equation for regions of inviscid flow, we use the vector identity →
→ →
→
→ → → V2 b % V * (§ * V ) 2
(V $ §)V ! § a
Show that this →vector identity is satisfied for→ the case of→ → velocity vector V in Cartesian coordinates, i.e., V ! ui # vj → # wk . For full credit, expand each term as far as possible and show all your work. 10–39 In the derivation of the Bernoulli equation for regions of inviscid flow, we rewrite the steady, incompressible Euler equation into a form showing that the gradient of three scalar terms is equal to the velocity vector crossed with the vorticity vector, noting that z is vertically upward → P → → V2 # gzb ! V * z §a # r 2
We then employ some arguments about the direction of the gradient vector and the direction of the cross product of two vectors to show that the sum of the three scalar terms must be constant along a streamline. In this problem you will use a different approach to achieve the same result. Namely, take the dot product of both sides of the Euler equation with
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554 FLUID MECHANICS →
velocity vector V and apply some fundamental rules about the dot product of two vectors. Sketches may be helpful. 10–40 Write out the components of the Euler equation as far as possible in Cartesian coordinates (x, y, z) and (u, v, w). Assume gravity acts in some arbitrary direction. 10–41 Write out the components of the Euler equation as far as possible in cylindrical coordinates (r, u, z) and (ur, uu, uz). Assume gravity acts in some arbitrary direction. 10–42 Water at T ! 20°C rotates as a rigid body about the z-axis in a spinning cylindrical container (Fig. P10–42). There are no viscous stresses since the water moves as a solid body; thus the Euler equation is appropriate. (We neglect viscous stresses caused by air acting on the water surface.) Integrate the Euler equation to generate an expression for pressure as a function of r and z everywhere in the water. Write an equation for the shape of the free surface (zsurface as a function of r). (Hint: P ! Patm everywhere on the free surface. The flow is rotationally symmetric about the z-axis.) Answer: zsurface ! v2r2/2g
r=R
r ur(r)
∆u
FIGURE P10–45 Sketch what the velocity profile at radius r would look like if friction were not neglected (i.e., a real flow) at the same volume flow rate. 10–46 In a certain region of steady, two-dimensional, → incompressible flow, the velocity field is given by V ! (u, v) → → ! (ax # b)i # (%ay # cx)j . Show that this region of flow can be considered inviscid.
Irrotational (Potential) Flow 10–47C What flow property determines whether a region of flow is rotational or irrotational? Discuss. v
Free surface P = Patm
10–48 In an irrotational region of flow, we can write the velocity vector as the gradient of the scalar velocity potential → → → function, V ! +f. The components of V in cylindrical coordinates, (r, u, z) and (ur, uu, uz), are
z
ur !
r
uu = vr ur = uz = 0
R
Water
FIGURE P10–42 10–43 Repeat Prob. 10–42, except let the rotating fluid be engine oil at 60°C. Discuss. 10–44 Using the results of Prob. 10–42, calculate the Bernoulli constant as a function of radial coordinate r. Answer:
Patm 2 2 r #v r
10–45 Consider steady, incompressible, two-dimensional flow of fluid into a converging duct # with straight walls (Fig. P10–45). The volume flow rate is V , and the velocity is in the radial direction only, with ur a function of r only. Let b be the width into the page. At the inlet into the converging duct (r ! R), ur ! ur(R). Assuming inviscid flow everywhere, generate an expression for ur as a function of r, R, and ur(R) only.
uu ! uz !
"f "r 1 "f r "u "f "z
From Chap. 9, we also write the components of the vor1 "u z "u u ticity vector in cylindrical coordinates as zr ! % , r "u "z "u r "u z 1 " 1 "u r . Substitute the zu ! % , and zz ! aru ub % r "r r "u "z "r velocity components into the vorticity components to show that all three components of the vorticity vector are indeed zero in an irrotational region of flow. 10–49 Substitute the components of the velocity vector given in Prob. 10–48 into the Laplace equation in cylindrical coordinates. Showing all your algebra, verify that the Laplace equation is valid in an irrotational region of flow. 10–50 Consider the flow field produced by a hair dryer (Fig. P10–50). Identify regions of this flow field that can be approximated as irrotational, and those for which the irrota-
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tional flow approximation would not be appropriate (rotational flow regions).
(potential) region of flow is not the same as an inviscid region of flow (Fig. P10–55). Discuss the differences and similarities between these two approximations. Give an example of each.
Inviscid? Irrotational? High Med low Off
?
FIGURE P10–50 10–51 Write the Bernoulli equation, and discuss how it differs between an inviscid, rotational region of flow and a viscous, irrotational region of flow. Which case is more restrictive (in regards to the Bernoulli equation)? 10–52 Streamlines in a steady, two-dimensional, incompressible flow field are sketched in Fig. P10–52. The flow in the region shown is also approximated as irrotational. Sketch what a few equipotential curves (curves of constant potential function) might look like in this flow field. Explain how you arrive at the curves you sketch.
Streamlines
FIGURE P10–55 10–56 Consider a steady, two-dimensional, incompressible, irrotational velocity field specified by its velocity potential function, f ! 5(x2 " y2) # 2x " 4y. (a) Calculate velocity components u and v. (b) Verify that the velocity field is irrotational in the region in which f applies. (c) Generate an expression for the stream function in this region. 10–57 Consider a planar irrotational region of flow in the ru-plane. Show that stream function c satisfies the Laplace equation in cylindrical coordinates. 10–58 In this chapter, we describe axisymmetric irrotational flow in terms of cylindrical coordinates r and z and velocity components ur and uz. An alternative description of axisymmetric flow arises if we use spherical polar coordinates and set the x-axis as the axis of symmetry. The two y or z
FIGURE P10–52 10–53 In an irrotational region of flow, the velocity field can be calculated without need of the momentum equation by solving the Laplace equation for velocity → potential function from the definif, and then solving→for the components of V → tion of f, namely, V ! $f. Discuss the role of the momentum equation in an irrotational region of flow. 10–54 Consider the following→ steady, two-dimensional,→ incompressible→ velocity field: V ! (u, v) ! (ax # b)i # ("ay # cx)j . Is this flow field irrotational? If so, generate an expression for the velocity potential function. Answers:
r
ur
u
Axisymmetric body
Yes, a(x 2 " y 2)/2 # bx # cy # constant
10–55 A subtle point, often missed by students of fluid mechanics (and even their professors!), is that an irrotational
uu
Rotational symmetry f
x
FIGURE P10–58
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relevant directional components are now r and u, and their corresponding velocity components are ur and uu. In this coordinate system, radial location r is the distance from the origin, and polar angle u is the angle of inclination between the radial vector and the axis of rotational symmetry (the xaxis), as sketched in Fig. P10–58; a slice defining the ruplane is shown. This is a type of two-dimensional flow because there are only two independent spatial variables, r and u. In other words, a solution of the velocity and pressure fields in any ru-plane is sufficient to characterize the entire region of axisymmetric irrotational flow. Write the Laplace equation for f in spherical polar coordinates, valid in regions of axisymmetric irrotational flow. (Hint: You may consult a textbook on vector analysis.) 10–59
Show that the incompressible continuity equation for
axisymmetric flow in spherical polar coordinates,
10–62
Consider an irrotational line vortex of strength / in "f the xy- or ru-plane. The velocity components are u r ! "r "c 1 "f 1 "c / ! ! 0 and u u ! !% ! . Generate expresr "u r "u "r 2pr sions for the velocity potential function and the stream function for the line vortex, showing all your work. 10–63 The stream function for steady, incompressible, twodimensional flow over a circular cylinder of radius a and free-stream velocity V& is c ! V& sin u(r % a2/r) for the case in which the flow field is approximated as irrotational (Fig. P10–63). Generate an expression for the velocity potential function f for this flow as a function of r and u, and parameters V& and a. y
1 " 2 (r u r) r "r
V∞
r
1 " # (u sin u) ! 0, is identically satisfied by a stream sin u "u u 1 "c 1 "c function defined as u r ! % 2 and u u ! , so r sin u "r r sin u "u long as c is a smooth function of r and u. 10–60 Consider a uniform stream of magnitude V inclined at angle a (Fig. P10–60). Assuming incompressible planar irrotational flow, find the velocity potential function and the stream function. Show all your work. Answers: f ! Vx cos a # Vy sin a, c ! Vy cos a % Vx sin a
y
V
a x
FIGURE P10–60 . Consider an irrotational line source of strength V/L "f in the xy- or ru-plane. The velocity components are u r ! "r # "c 1 "c V ,L 1 "f and u u ! ! ! ! % ! 0. In this chapter, r "u 2pr r "u "r we started with the equation for uu to generate expressions for the velocity potential function and the stream function for the line source. Repeat the analysis, except start with the equation for ur, showing all your work.
a
u x
FIGURE P10–63 10–64 What is D’Alembert’s paradox? Why is it a paradox?
Boundary Layers 10–65C In this chapter, we make a statement that the boundary layer approximation “bridges the gap” between the Euler equation and the Navier–Stokes equation. Explain. 10–66C For each statement, choose whether the statement is true or false and discuss your answer briefly. These statements concern a laminar boundary layer on a flat plate (Fig. P10–66C). (a) At a given x-location, if the Reynolds number were to increase, the boundary layer thickness would also increase. (b) As outer flow velocity increases, so does the boundary layer thickness. (c) As the fluid viscosity increases, so does the boundary layer thickness. (d) As the fluid density increases, so does the boundary layer thickness. y
10–61
U(x) = V
Outer flow x
V
Boundary layer
FIGURE P10–66C
d(x)
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557 CHAPTER 10 V
U(x) = V
y
d(x) x
Boundary layer
FIGURE P10–68C 10–67C We usually think of boundary layers as occurring along solid walls. However, there are other flow situations in which the boundary layer approximation is also appropriate. Name three such flows, and explain why the boundary layer approximation is appropriate. 10–68C A laminar boundary layer growing along a flat plate is sketched in Fig. P10–68C. Several velocity profiles and the boundary layer thickness d(x) are also shown. Sketch several streamlines in this flow field. Is the curve representing d(x) a streamline? 10–69C What is a trip wire, and what is its purpose? 10–70 Air at 30°C flows at a uniform speed of 25.0 m/s along a smooth flat plate. Calculate the approximate x-location along the plate where the boundary layer begins the transition process toward turbulence. At approximately what xlocation along the plate is the boundary layer likely to be fully turbulent? Answers: 6 to 7 cm, 2 m 10–71E Water flows over the fin of a small underwater vehicle at a speed of V ! 6.0 mi/h (Fig. P10–71E). The temperature of the water is 40°F, and the chord length c of the fin is 1.6 ft. Is the boundary layer on the surface of the fin laminar or turbulent or transitional? V
Boundary layer
80°F, and the length of the wind tunnel test section is 1.5 ft. Assume that the boundary layer thickness is negligible prior to the start of the test section. Is the boundary layer along the test section wall laminar or turbulent or transitional? Answer: laminar
10–74 Static pressure P is measured at two locations along the wall of a laminar boundary layer (Fig. P10–74). The measured pressures are P1 and P2, and the distance between the taps is small compared to the characteristic body dimension (1x ! x2 % x1 '' L). The outer flow velocity above the boundary layer at point 1 is U1. The fluid density and viscosity are r and m, respectively. Generate an approximate expression for U2, the outer flow velocity above the boundary layer at point 2, in terms of P1, P2, 1x, U1, r, and m.
U2
Outer flow
Boundary layer
U1 d P1 x1
x
P2 x2
Wall Pressure taps P1
P2
FIGURE P10–74 c
FIGURE P10–71E 10–72 Air flows parallel to a speed limit sign along the highway at speed V ! 5.0 m/s. The temperature of the air is 25°C, and the width W of the sign parallel to the flow direction is 0.45 m. Is the boundary layer on the sign laminar or turbulent or transitional? 10–73E Air flows through the test section of a small wind tunnel at speed V ! 7.5 ft/s. The temperature of the air is
10–75 Consider two pressure taps along the wall of a laminar boundary layer as in Fig. P10–74. The fluid is air at 25°C, U1 ! 10.3 m/s, and the static pressure P1 is 2.44 Pa greater than static pressure P2, as measured by a very sensitive differential pressure transducer. Is outer flow velocity U2 greater than, equal to, or less than outer flow velocity U1? Explain. Estimate U2. Answers: Less than, 10.0 m/s 10–76 In your own words, summarize the five steps of the boundary layer procedure. 10–77 In your own words, list at least three “red flags” to look out for when performing laminar boundary layer calculations.
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10–78 Consider the Blasius solution for a laminar flat plate boundary layer. The nondimensional slope at the wall is given by Eq. 8 of Example 10–10. Transform this result to physical variables, and show that Eq. 9 of Example 10–10 is correct. 10–79E For the small wind tunnel of Prob. 10–73E, assume the flow remains laminar, and estimate the boundary layer thickness, the displacement thickness, and the momentum thickness of the boundary layer at the end of the test section. Give your answers in inches, compare the three results, and discuss. 10–80 One dimension of a rectangular flat plate is twice the other. Air at uniform speed flows parallel to the plate, and a laminar boundary layer forms on both sides of the plate. Which orientation—long dimension to the wind (Fig. P10–80a) or short dimension to the wind (Fig. P10–80b)— has the higher drag? Explain.
V
(a)
V
(b)
FIGURE P10–80
V
x
FIGURE P10–84 ary layer in a large wind tunnel (Fig. P10–84). The scoop is constructed of thin sheet metal. The air is at 20°C, and flows at V ! 65.0 m/s. How high (dimension h) should the scoop be at downstream distance x ! 1.45 m? 10–85E A small, axisymmetric, low-speed wind tunnel is built to calibrate hot wires. The diameter of the test section is 6.0 in, and its length is 10.0 in. The air is at 70°F. At a uniform air speed of 5.0 ft/s at the test section inlet, by how much will the centerline air speed accelerate by the end of the test section? What should the engineers do to eliminate this acceleration? 10–86E Air at 70°F flows parallel to a smooth, thin, flat plate at 15.5 ft/s. The plate is 10.6 ft long. Determine whether the boundary layer on the plate is most likely laminar, turbulent, or somewhere in between (transitional). Compare the boundary layer thickness at the end of the plate for two cases: (a) the boundary layer is laminar everywhere, and (b) the boundary layer is turbulent everywhere. Discuss. 10–87 Air at 20°C flows at V ! 5.0 m/s parallel to a flat plate (Fig. P10–87). The front of the plate is well rounded, and the plate is 40 cm long. The plate thickness is h ! 0.75 cm, but because of boundary layer displacement effects, the flow outside the boundary layer “sees” a plate that has larger apparent thickness. Calculate the apparent thickness of the plate (include both sides) at downstream distance x ! 25 cm. Answer: 3.75 cm
10–81 Two definitions of displacement thickness are given in this chapter. Write both definitions in your own words. For the laminar boundary layer growing on a flat plate, which is larger—boundary layer thickness d or displacement thickness d*? Discuss. 10–82 A laminar flow wind tunnel has a test section that is 40 cm in diameter and 60 cm in length. The air is at 20°C. At a uniform air speed of 2.0 m/s at the test section inlet, by how much will the centerline air speed accelerate by the end of the test section? Answer: Approx. 4% 10–83 Repeat the calculation of Prob. 10–82, except for a test section of square rather than round cross section, with a 40 cm * 40 cm cross section and a length of 60 cm. Compare the result to that of Prob. 10–82 and discuss. 10–84 In order to avoid boundary layer interference, engineers design a “boundary layer scoop” to skim off the bound-
h
V
h x
FIGURE P10–87 10–88
Air at 20°C flows at V ! 80.0 m/s over a smooth flat plate of length L ! 17.5 m. Plot the turbulent boundary layer profile in physical variables (u as a function of y) at x ! L. Compare the profile generated by the one-seventh-power law, the log law, and Spalding’s law of the wall, assuming that the boundary layer is fully turbulent from the beginning of the plate. 10–89 Explain the difference between a favorable and an adverse pressure gradient in a boundary layer. In which case does the pressure increase downstream? Why?
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559 CHAPTER 10
10–90 Discuss the implication of an inflection point in a boundary layer profile. Specifically, does the existence of an inflection point infer a favorable or adverse pressure gradient? Explain. 10–91 Compare flow separation for a laminar versus turbulent boundary layer. Specifically, which case is more resistant to flow separation? Why? Based on your answer, explain why golf balls have dimples. 10–92 The streamwise velocity component of a steady, incompressible, laminar, flat plate boundary layer of boundary layer thickness d is approximated by the simple linear expression, u ! Uy/d for y ' d, and u ! U for y . d (Fig. P10–92). Generate expressions for displacement thickness and momentum thickness as functions of d, based on this linear approximation. Compare the approximate values of d*/d and u/d to the values of d*/d and u/d obtained from the Blasius solution. Answers: 0.500, 0.167 U(x) = V
V
10–95 The streamwise velocity component of a steady, incompressible, laminar, flat plate boundary layer of boundary layer thickness d is approximated by the sine wave profile of Prob. 10–94. Generate expressions for displacement thickness and momentum thickness as functions of d, based on this sine wave approximation. Compare the approximate values of d*/d and u/d to the values of d*/d and u/d obtained from the Blasius solution. 10–96 For the sine wave approximation of Prob. 10–94, use the definition of local skin friction coefficient and the Kármán integral equation to generate an expression for d/x. Compare your result to the Blasius expression for d/x. (Note: You will also need the results of Prob. 10–95 to do this problem.) 10–97 Compare shape factor H (defined in Eq. 10–95) for a laminar versus a turbulent boundary layer on a flat plate, assuming that the turbulent boundary layer is turbulent from the beginning of the plate. Discuss. Specifically, why do you suppose H is called a “shape factor”? Answers: 2.59, 1.25 to 1.30
10–98 Calculate the value of shape factor H for the limiting case of a boundary layer that is infinitesimally thin (Fig. P10–98). This value of H is the minimum possible value.
d(x)
10–99 Integrate Eq. 5 to obtain Eq. 6 of Example 10–14, showing all your work.
x
FIGURE P10–92 10–93 For the linear approximation of Prob. 10–92, use the definition of local skin friction coefficient and the Kármán integral equation to generate an expression for d/x. Compare your result to the Blasius expression for d/x. (Note: You will need the results of Prob. 10–92 to do this problem.) 10–94
The Blasius boundary layer profile is an exact solution of the boundary layer equations for flow over a flat plate. However, the results are somewhat cumbersome to use, since the data appear in tabular form (the solution is numerical). Thus, a simple sine wave approximation (Fig. P10–94) is often used in place of the Blasius solupy tion, namely, u(y) " U sin a b for y ' d, and u ! U for 2d y '' d, where d is the boundary layer thickness. Plot the Blasius profile and the sine wave approximation on the same plot, in nondimensional form (u/U versus y/d), and compare. Is the sine wave profile a reasonable approximation? U(x) = V
V d(x) x
FIGURE P10–94
U(x)
x
FIGURE P10–98 10–100 Consider a turbulent boundary layer on a flat plate. Suppose only two things are known: Cf, x ≅ 0.059 · (Rex)%1/5 and u ≅ 0.097d. Use the Kármán integral equation to generate an expression for d/x, and compare your result to column (b) of Table 10–4.
Review Problems 10–101C For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) The velocity potential function can be defined for threedimensional flows. (b) The vorticity must be zero in order for the stream function to be defined. (c) The vorticity must be zero in order for the velocity potential function to be defined. (d) The stream function can be defined only for two-dimensional flow fields.
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560 FLUID MECHANICS
10–102 In this chapter, we discuss solid body rotation (Fig. P10–102) as an example of an inviscid flow that is also rotational. The velocity components are ur ! 0, uu ! vr, and uz ! 0. Compute the viscous term of the u-component of the Navier–Stokes equation, and discuss. Verify that this velocity field is indeed rotational by computing the z-component of vorticity. Answer: zz ! 2v
uu uu = vr
r
velocity components are ur ! 0, uu ! //(2pr), and uz ! 0. Compute the viscous term of the u-component of the Navier–Stokes equation, and discuss. Verify that this velocity field is indeed irrotational by computing the z-component of vorticity. 10–105 Calculate the nine components of the viscous stress tensor in cylindrical coordinates (see Chap. 9) for the velocity field of Prob. 10–104. Discuss. 10–106 Water falls down a vertical pipe by gravity alone. The flow between vertical locations z1 and z2 is fully developed, and velocity profiles at these two locations are sketched in Fig. P10–106. Since there is no forced pressure gradient, pressure P is constant everywhere in the flow (P ! Patm). Calculate the modified pressure at locations z1 and z2. Sketch profiles of modified pressure at locations z1 and z2. Discuss.
z = z2
FIGURE P10–102 10–103 Calculate the nine components of the viscous stress tensor in cylindrical coordinates (see Chap. 9) for the velocity field of Prob. 10–102. Discuss.
z = z1
10–104 In this chapter, we discuss the line vortex (Fig. P10–104) as an example of an irrotational flow field. The
→
g z
FIGURE P10–106 uu
L
uu = 2pr
r
10–107 Suppose the vertical pipe of Prob. 10–106 is now horizontal instead. In order to achieve the same volume flow rate as that of Prob. 10–106, we must supply a forced pressure gradient. Calculate the required pressure drop between two axial locations in the pipe that are the same distance apart as z2 and z1 of Fig. P10–106. How does modified pressure P( change between the vertical and horizontal cases?
Design and Essay Problem
FIGURE P10–104
10–108 Explain why there is a significant velocity overshoot for the midrange values of the Reynolds number in the velocity profiles of Fig. 10–136, but not for the very small values of Re or for the very large values of Re.
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CHAPTER
11
FLOW OVER BODIES: DRAG AND LIFT
I
n Chap. 8, we considered the flow of fluids inside pipes, with emphasis on pressure drop and head losses and their relations to flow rate. In this chapter, we consider the flow of fluids over bodies that are immersed in a fluid, called external flow, with emphasis on the resulting lift and drag forces. External flow is characterized by a freely growing boundary layer surrounded by an outer flow region that involves small velocity and temperature gradients. In internal flows, the entire flow field is dominated by viscous effects, while in external flow, the viscous effects are confined to a portion of the flow field such as the boundary layers and wakes. When a fluid moves over a solid body, it exerts pressure forces normal to the surface and shear forces parallel to the surface along the outer surface of the body. We are usually interested in the resultant of the pressure and shear forces acting on the body rather than the details of the distributions of these forces along the entire surface of the body. The component of the resultant pressure and shear forces that acts in the flow direction is called the drag force (or just drag), and the component that acts normal to the flow direction is called the lift force (or just lift). We start this chapter with a discussion of drag and lift, and explore the concepts of pressure drag, friction drag, and flow separation. We continue with the drag coefficients of various two- and three-dimensional geometries encountered in practice and determine the drag force using experimentally determined drag coefficients. We then examine the development of the velocity boundary layer during parallel flow over a flat surface, and develop relations for the skin friction and drag coefficients for flow over flat plates, cylinders, and spheres. Finally, we discuss the lift developed by airfoils and the factors that affect the lift characteristics of bodies.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
■
Have an intuitive understanding of the various physical phenomena such as drag, friction and pressure drag, drag reduction, and lift Calculate the drag force associated with flow over common geometries Understand the effects of flow regime on the drag coefficients associated with flow over cylinders and spheres Understand the fundamentals of flow over airfoils, and calculate the drag and lift forces acting on airfoils
561
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562 FLUID MECHANICS
11–1
FIGURE 11–1 Flow over bodies is commonly encountered in practice.
■
INTRODUCTION
Fluid flow over solid bodies frequently occurs in practice, and it is responsible for numerous physical phenomena such as the drag force acting on automobiles, power lines, trees, and underwater pipelines; the lift developed by airplane wings; upward draft of rain, snow, hail, and dust particles in high winds; the transportation of red blood cells by blood flow; the entrainment and disbursement of liquid droplets by sprays; the vibration and noise generated by bodies moving in a fluid; and the power generated by wind turbines (Fig. 11–1). Therefore, developing a good understanding of external flow is important in the design of many engineering systems such as aircraft, automobiles, buildings, ships, submarines, and all kinds of turbines. Late-model cars, for example, have been designed with particular emphasis on aerodynamics. This has resulted in significant reductions in fuel consumption and noise, and considerable improvement in handling. Sometimes a fluid moves over a stationary body (such as the wind blowing over a building), and other times a body moves through a quiescent fluid (such as a car moving through air). These two seemingly different processes are equivalent to each other; what matters is the relative motion between the fluid and the body. Such motions are conveniently analyzed by fixing the coordinate system on the body and are referred to as flow over bodies or external flow. The aerodynamic aspects of different airplane wing designs, for example, are studied conveniently in a lab by placing the wings in a wind tunnel and blowing air over them by large fans. Also, a flow can be classified as being steady or unsteady, depending on the reference frame selected. Flow around an airplane, for example, is always unsteady with respect to the ground, but it is steady with respect to a frame of reference moving with the airplane at cruise conditions. The flow fields and geometries for most external flow problems are too complicated to be solved analytically, and thus we have to rely on correlations based on experimental data. The availability of high-speed computers has made it possible to conduct series of “numerical experiments” quickly by solving the governing equations numerically (Chap. 15), and to resort to the expensive and time-consuming testing and experimentation only in the final stages of design. Such testing is done in wind tunnels. H. F. Phillips (1845–1912) built the first wind tunnel in 1894 and measured lift and drag. In this chapter we mostly rely on relations developed experimentally. The velocity of the fluid approaching a body is called the free-stream velocity and is denoted by V. It is also denoted by u! or U! when the flow is aligned with the x-axis since u is used to denote the x-component of velocity. The fluid velocity ranges from zero at the surface (the no-slip condition) to the free-stream value away from the surface, and the subscript “infinity” serves as a reminder that this is the value at a distance where the presence of the body is not felt. The free-stream velocity may vary with location and time (e.g., the wind blowing past a building). But in the design and analysis, the free-stream velocity is usually assumed to be uniform and steady for convenience, and this is what we will do in this chapter. The shape of a body has a profound influence on the flow over the body and the velocity field. The flow over a body is said to be two-dimensional when the body is very long and of constant cross section and the flow is
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563 CHAPTER 11
normal to the body. The wind blowing over a long pipe perpendicular to its axis is an example of two-dimensional flow. Note that the velocity component in the axial direction is zero in this case, and thus the velocity is two-dimensional. The two-dimensional idealization is appropriate when the body is sufficiently long so that the end effects are negligible and the approach flow is uniform. Another simplification occurs when the body possesses rotational symmetry about an axis in the flow direction. The flow in this case is also two-dimensional and is said to be axisymmetric. A bullet piercing through air is an example of axisymmetric flow. The velocity in this case varies with the axial distance x and the radial distance r. Flow over a body that cannot be modeled as two-dimensional or axisymmetric, such as flow over a car, is three-dimensional (Fig. 11–2). Flow over bodies can also be classified as incompressible flows (e.g., flows over automobiles, submarines, and buildings) and compressible flows (e.g., flows over high-speed aircraft, rockets, and missiles). Compressibility effects are negligible at low velocities (flows with Ma " 0.3), and such flows can be treated as incompressible. Compressible flow is discussed in Chap. 12, and flows that involve partially immersed bodies with a free surface (such as a ship cruising in water) are beyond the scope of this introductory text. Bodies subjected to fluid flow are classified as being streamlined or blunt, depending on their overall shape. A body is said to be streamlined if a conscious effort is made to align its shape with the anticipated streamlines in the flow. Streamlined bodies such as race cars and airplanes appear to be contoured and sleek. Otherwise, a body (such as a building) tends to block the flow and is said to be bluff or blunt. Usually it is much easier to force a streamlined body through a fluid, and thus streamlining has been of great importance in the design of vehicles and airplanes (Fig. 11–3).
11–2
■
Long cylinder (2-D) r x Bullet (axisymmetric)
Car (3-D)
FIGURE 11–2 Two-dimensional, axisymmetric, and three-dimensional flows.
DRAG AND LIFT
It is a common experience that a body meets some resistance when it is forced to move through a fluid, especially a liquid. As you may have noticed, it is very difficult to walk in water because of the much greater resistance it offers to motion compared to air. Also, you may have seen high winds knocking down trees, power lines, and even trailers and felt the strong “push” the wind exerts on your body (Fig. 11–4). You experience the same feeling when you extend your arm out of the window of a moving car. A fluid may exert forces and moments on a body in and about various directions. The force a flowing fluid exerts on a body in the flow direction is called drag. The drag force can be measured directly by simply attaching the body subjected to fluid flow to a calibrated spring and measuring the displacement in the flow direction (just like measuring weight with a spring scale). More sophisticated drag-measuring devices, called drag balances, use flexible beams fitted with strain gages to measure the drag electronically. Drag is usually an undesirable effect, like friction, and we do our best to minimize it. Reduction of drag is closely associated with the reduction of fuel consumption in automobiles, submarines, and aircraft; improved safety and durability of structures subjected to high winds; and reduction of noise
60 mi/h
70 hp
60 mi/h
50 hp
FIGURE 11–3 It is much easier to force a streamlined body than a blunt body through a fluid.
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564 FLUID MECHANICS
FIGURE 11–4 High winds knock down trees, power lines, and even people as a result of the drag force. FL
FR f
FD
FD = FR cos f FL = FR sin f
Outer normal → n P dA u tw dA
A
P (absolute)
tw
FIGURE 11–5 The pressure and viscous forces acting on a two-dimensional body and the resultant lift and drag forces.
and vibration. But in some cases drag produces a very beneficial effect and we try to maximize it. Friction, for example, is a “life saver” in the brakes of automobiles. Likewise, it is the drag that makes it possible for people to parachute, for pollens to fly to distant locations, and for us all to enjoy the waves of the oceans and the relaxing movements of the leaves of trees. A stationary fluid exerts only normal pressure forces on the surface of a body immersed in it. A moving fluid, however, also exerts tangential shear forces on the surface because of the no-slip condition caused by viscous effects. Both of these forces, in general, have components in the direction of flow, and thus the drag force is due to the combined effects of pressure and wall shear forces in the flow direction. The components of the pressure and wall shear forces in the direction normal to the flow tend to move the body in that direction, and their sum is called lift. For two-dimensional flows, the resultant of the pressure and shear forces can be split into two components: one in the direction of flow, which is the drag force, and another in the direction normal to flow, which is the lift, as shown in Fig. 11–5. For three-dimensional flows, there is also a side force component in the direction normal to the page that tends to move the body in that direction. The fluid forces also may generate moments and cause the body to rotate. The moment about the flow direction is called the rolling moment, the moment about the lift direction is called the yawing moment, and the moment about the side force direction is called the pitching moment. For bodies that possess symmetry about the lift–drag plane such as cars, airplanes, and ships, the side force, the yawing moment, and the rolling moment are zero when the wind and wave forces are aligned with the body. What remain for such bodies are the drag and lift forces and the pitching moment. For axisymmetric bodies aligned with the flow, such as a bullet, the only force exerted by the fluid on the body is the drag force. The pressure and shear forces acting on a differential area dA on the surface are PdA and tw dA, respectively. The differential drag force and the lift force acting on dA in two-dimensional flow are (Fig. 11–5) dFD # $P dA cos u % tw dA sin u
(11–1)
dFL # $P dA sin u $ tw dA cos u
(11–2)
and
where u is the angle that the outer normal of dA makes with the positive flow direction. The total drag and lift forces acting on the body are determined by integrating Eqs. 11–1 and 11–2 over the entire surface of the body, Drag force:
FD #
! dF # ! ($P cos u % t
sin u) dA
(11–3)
! dF # $ ! (P sin u % t
cos u) dA
(11–4)
D
A
w
A
and Lift force:
FL #
L
A
w
A
These are the equations used to predict the net drag and lift forces on bodies when the flow is simulated on a computer (Chap. 15). However, when we perform experimental analyses, Eqs. 11–3 and 11–4 are not practical since
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565 CHAPTER 11
the detailed distributions of pressure and shear forces are difficult to obtain by measurements. Fortunately, this information is often not needed. Usually all we need to know is the resultant drag force and lift acting on the entire body, which can be measured directly and easily in a wind tunnel. Equations 11–1 and 11–2 show that both the skin friction (wall shear) and pressure, in general, contribute to the drag and the lift. In the special case of a thin flat plate aligned parallel to the flow direction, the drag force depends on the wall shear only and is independent of pressure since u # 90°. When the flat plate is placed normal to the flow direction, however, the drag force depends on the pressure only and is independent of wall shear since the shear stress in this case acts in the direction normal to flow and u # 0° (Fig. 11–6). If the flat plate is tilted at an angle relative to the flow direction, then the drag force depends on both the pressure and the shear stress. The wings of airplanes are shaped and positioned specifically to generate lift with minimal drag. This is done by maintaining an angle of attack during cruising, as shown in Fig. 11–7. Both lift and drag are strong functions of the angle of attack, as we discuss later in this chapter. The pressure difference between the top and bottom surfaces of the wing generates an upward force that tends to lift the wing and thus the airplane to which it is connected. For slender bodies such as wings, the shear force acts nearly parallel to the flow direction, and thus its contribution to the lift is small. The drag force for such slender bodies is mostly due to shear forces (the skin friction). The drag and lift forces depend on the density r of the fluid, the upstream velocity V, and the size, shape, and orientation of the body, among other things, and it is not practical to list these forces for a variety of situations. Instead, it is found convenient to work with appropriate dimensionless numbers that represent the drag and lift characteristics of the body. These numbers are the drag coefficient CD, and the lift coefficient CL, and they are defined as Drag coefficient:
CD # 1
Lift coefficient:
CL # 1
FD
2 rV
2
A
FL
2 2 rV A
Boundary layer
y
u
where A is ordinarily the frontal area (the area projected on a plane normal to the direction of flow) of the body. In other words, A is the area that would be seen by a person looking at the body from the direction of the approaching fluid. The frontal area of a cylinder of diameter D and length L, for example, is A # LD. In lift calculations of some thin bodies, such as airfoils, A is taken to be the planform area, which is the area seen by a person looking at the body from above in a direction normal to the body. The drag and lift coefficients are primarily functions of the shape of the body, but in some cases they also depend on the Reynolds number and the surface roughness. The term 12 rV 2 in Eqs. 11–5 and 11–6 is the dynamic pressure. The local drag and lift coefficients vary along the surface as a result of the changes in the velocity boundary layer in the flow direction. We are usually interested in the drag and lift forces for the entire surface, which can be determined using the average drag and lift coefficients. Therefore, we present correlations for both local (identified with the subscript x) and average drag
tw
tw
(a) High pressure + + + + + + + +
Low pressure
– – – – – – – – Wall shear (b)
FIGURE 11–6 (a) Drag force acting on a flat plate parallel to the flow depends on wall shear only. (b) Drag force acting on a flat plate normal to the flow depends on the pressure only and is independent of the wall shear, which acts normal to the free-stream flow.
Lift
(11–5)
(11–6)
x
tw
V Drag
FIGURE 11–7 Airplane wings are shaped and positioned to generate sufficient lift during flight while keeping drag at a minimum. Pressures above and below atmospheric pressure are indicated by plus and minus signs, respectively.
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566 FLUID MECHANICS FB
Vterminal
and lift coefficients. When relations for local drag and lift coefficients for a surface of length L are available, the average drag and lift coefficients for the entire surface can be determined by integration from
W FD
CD #
CL #
FD
FIGURE 11–9 Schematic for Example 11–1.
1 L
!
L
C D, x dx
(11–7)
C L, x dx
(11–8)
0
0
When a body is dropped into the atmosphere or a lake, it first accelerates under the influence of its weight. The motion of the body is resisted by the drag force, which acts in the direction opposite to motion. As the velocity of the body increases, so does the drag force. This continues until all the forces balance each other and the net force acting on the body (and thus its acceleration) is zero. Then the velocity of the body remains constant during the rest of its fall if the properties of the fluid in the path of the body remain essentially constant. This is the maximum velocity a falling body can attain and is called the terminal velocity (Fig. 11–8). The forces acting on a falling body are usually the drag force, the buoyant force, and the weight of the body. EXAMPLE 11–1
Wind tunnel 60 mi/h
!
L
and
FD = W – FB (No acceleration)
FIGURE 11–8 During a free fall, a body reaches its terminal velocity when the drag force equals the weight of the body minus the buoyant force.
1 L
Measuring the Drag Coefficient of a Car
The drag coefficient of a car at the design conditions of 1 atm, 70°F, and 60 mi/h is to be determined experimentally in a large wind tunnel in a fullscale test (Fig. 11–9). The frontal area of the car is 22.26 ft2. If the force acting on the car in the flow direction is measured to be 68 lbf, determine the drag coefficient of this car.
SOLUTION The drag force acting on a car is measured in a wind tunnel. The drag coefficient of the car at test conditions is to be determined. Assumptions 1 The flow of air is steady and incompressible. 2 The cross section of the tunnel is large enough to simulate free flow over the car. 3 The bottom of the tunnel is also moving at the speed of air to approximate actual driving conditions or this effect is negligible. Properties The density of air at 1 atm and 70°F is r # 0.07489 lbm/ft3. Analysis The drag force acting on a body and the drag coefficient are given by FD # C D A
rV 2 2
and
CD #
2FD rAV 2
where A is the frontal area. Substituting and noting that 1 mi/h # 1.467 ft/s, the drag coefficient of the car is determined to be
CD #
2 & (68 lbf) 32.2 lbm ' ft/s2 a b # 0.34 2 2 1 lbf (0.07489 lbm/ft )(22.26 ft )(60 & 1.467 ft/s) 3
Discussion Note that the drag coefficient depends on the design conditions, and its value may be different at different conditions such as the Reynolds number. Therefore, the published drag coefficients of different vehicles can be compared meaningfully only if they are determined under similar conditions. This shows the importance of developing standard testing procedures in industry.
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11–3
■
FRICTION AND PRESSURE DRAG
As mentioned in Section 11–2, the drag force is the net force exerted by a fluid on a body in the direction of flow due to the combined effects of wall shear and pressure forces. It is often instructive to separate the two effects, and study them separately. The part of drag that is due directly to wall shear stress tw is called the skin friction drag (or just friction drag FD, friction) since it is caused by frictional effects, and the part that is due directly to pressure P is called the pressure drag (also called the form drag because of its strong dependence on the form or shape of the body). The friction and pressure drag coefficients are defined as C D, friction #
FD, friction 1 2 2 rV A
and
C D, pressure #
FD, pressure 1 2 2 rV A
(11–9)
When the friction and pressure drag coefficients or forces are available, the total drag coefficient or drag force can be determined by simply adding them, C D # C D, friction % C D, pressure
and
FD # FD, friction % FD, pressure (11–10)
The friction drag is the component of the wall shear force in the direction of flow, and thus it depends on the orientation of the body as well as the magnitude of the wall shear stress tw. The friction drag is zero for a flat surface normal to flow, and maximum for a flat surface parallel to flow since the friction drag in this case equals the total shear force on the surface. Therefore, for parallel flow over a flat surface, the drag coefficient is equal to the friction drag coefficient, or simply the friction coefficient. Friction drag is a strong function of viscosity, and increases with increasing viscosity. The Reynolds number is inversely proportional to the viscosity of the fluid. Therefore, the contribution of friction drag to total drag for blunt bodies is less at higher Reynolds numbers and may be negligible at very high Reynolds numbers. The drag in such cases is mostly due to pressure drag. At low Reynolds numbers, most drag is due to friction drag. This is especially the case for highly streamlined bodies such as airfoils. The friction drag is also proportional to the surface area. Therefore, bodies with a larger surface area experience a larger friction drag. Large commercial airplanes, for example, reduce their total surface area and thus their drag by retracting their wing extensions when they reach cruising altitudes to save fuel. The friction drag coefficient is independent of surface roughness in laminar flow, but is a strong function of surface roughness in turbulent flow due to surface roughness elements protruding further into the boundary layer. The friction drag coefficient is analogous to the friction factor in pipe flow discussed in Chap. 8, and its value depends on the flow regime. The pressure drag is proportional to the frontal area and to the difference between the pressures acting on the front and back of the immersed body. Therefore, the pressure drag is usually dominant for blunt bodies, small for streamlined bodies such as airfoils, and zero for thin flat plates parallel to the flow (Fig. 11–10). The pressure drag becomes most significant when the velocity of the fluid is too high for the fluid to be able to follow the curvature of the body, and thus the fluid separates from the body at some point
FIGURE 11–10 Drag is due entirely to friction drag for a flat plate parallel to the flow; it is due entirely to pressure drag for a flat plate normal to the flow; and it is due to both (but mostly pressure drag) for a cylinder normal to the flow. The total drag coefficient CD is lowest for a parallel flat plate, highest for a vertical flat plate, and in between (but close to that of a vertical flat plate) for a cylinder. From G. M. Homsy, et al. (2000).
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and creates a very low pressure region in the back. The pressure drag in this case is due to the large pressure difference between the front and back sides of the body.
Reducing Drag by Streamlining
The first thought that comes to mind to reduce drag is to streamline a body in order to reduce flow separation and thus to reduce pressure drag. Even car salespeople are quick to point out the low drag coefficients of their cars, owing to streamlining. But streamlining has opposite effects on pressure and friction drags. It decreases pressure drag by delaying boundary layer separation and thus reducing the pressure difference between the front and back of the body and increases the friction drag by increasing the surface area. The end result depends on which effect dominates. Therefore, any optimization study to reduce the drag of a body must consider both effects and must attempt to minimize the sum of the two, as shown in Fig. 11–11. The minimum total drag occurs at D/L # 0.25 for the case shown in Fig. 11–11. For the case of a circular cylinder with the same thickness as the streamlined shape of Fig. 11–11, the drag coefficient would be about five times as much. Therefore, it is possible to reduce the drag of a cylindrical component to one-fifth by the use of proper fairings. The effect of streamlining on the drag coefficient can be described best by considering long elliptical cylinders with different aspect (or length-tothickness) ratios L/D, where L is the length in the flow direction and D is the thickness, as shown in Fig. 11–12. Note that the drag coefficient decreases drastically as the ellipse becomes slimmer. For the special case of L/D # 1 (a circular cylinder), the drag coefficient is CD " 1 at this Reynolds number. As the aspect ratio is decreased and the cylinder resembles a flat plate, the drag coefficient increases to 1.9, the value for a flat plate normal to flow. Note that the curve becomes nearly flat for aspect ratios greater than about 4. Therefore, for a given diameter D, elliptical shapes with an aspect ratio of about L/D " 4 usually offer a good compro-
CD = 1 2
0.12
FIGURE 11–11 The variation of friction, pressure, and total drag coefficients of a streamlined strut with thickness-to-chord length ratio for Re # 4 & 104. Note that CD for airfoils and other thin bodies is based on planform area rather than frontal area. From Abbott and von Doenhoff (1959).
V
FD
D L
rV 2LD
L = length
0.10
Total drag
0.08 0.06
Friction drag
0.04 0.02 0
Pressure drag 0
0.1
0.2 D/L
0.3
0.4
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569 CHAPTER 11 2.5 Flat plate normal to flow 2.0
VD Re = v = 10 5 L = length
D
V 1.5
L Circular cylinder
CD
mise between the total drag coefficient and length L. The reduction in the drag coefficient at high aspect ratios is primarily due to the boundary layer staying attached to the surface longer and the resulting pressure recovery. The friction drag on an elliptical cylinder with an aspect ratio of 4 is negligible (less than 2 percent of total drag at this Reynolds number). As the aspect ratio of an elliptical cylinder is increased by flattening it (i.e., decreasing D while holding L constant), the drag coefficient starts increasing and tends to infinity as L/D → ! (i.e., as the ellipse resembles a flat plate parallel to flow). This is due to the frontal area, which appears in the denominator in the definition of CD, approaching zero. It does not mean that the drag force increases drastically (actually, the drag force decreases) as the body becomes flat. This shows that the frontal area is inappropriate for use in the drag force relations for slim bodies such as thin airfoils and flat plates. In such cases, the drag coefficient is defined on the basis of the planform area, which is simply the surface area for a flat plate parallel to flow. This is quite appropriate since for slim bodies the drag is almost entirely due to friction drag, which is proportional to the surface area. Streamlining has the added benefit of reducing vibration and noise. Streamlining should be considered only for blunt bodies that are subjected to high-velocity fluid flow (and thus high Reynolds numbers) for which flow separation is a real possibility. It is not necessary for bodies that typically involve low Reynolds number flows (e.g., creeping flows in which Re ( 1) as discussed in Chap. 10, since the drag in those cases is almost entirely due to friction drag, and streamlining will only increase the surface area and thus the total drag. Therefore, careless streamlining may actually increase drag instead of decreasing it.
1.0 CD = 0.5
0
0
1
2
FD 1 rV 2bD 2
3 L /D
4
5
6
FIGURE 11–12 The variation of the drag coefficient of a long elliptical cylinder with aspect ratio. Here CD is based on the frontal area bD where b is the width of the body. From Blevins (1984).
Separation point
Flow Separation
When driving on country roads, it is a common safety measure to slow down at sharp turns in order to avoid being thrown off the road. Many drivers have learned the hard way that a car will refuse to comply when forced to turn curves at excessive speeds. We can view this phenomenon as “the separation of cars” from roads. This phenomenon is also observed when fast vehicles jump off hills. At low velocities, the wheels of the vehicle always remain in contact with the road surface. But at high velocities, the vehicle is too fast to follow the curvature of the road and takes off at the hill, losing contact with the road. A fluid acts much the same way when forced to flow over a curved surface at high velocities. A fluid climbs the uphill portion of the curved surface with no problem, but it has difficulty remaining attached to the surface on the downhill side. At sufficiently high velocities, the fluid stream detaches itself from the surface of the body. This is called flow separation (Fig. 11–13). Flow can separate from a surface even if it is fully submerged in a liquid or immersed in a gas (Fig. 11–14). The location of the separation point depends on several factors such as the Reynolds number, the surface roughness, and the level of fluctuations in the free stream, and it is usually difficult to predict exactly where separation will occur unless there are sharp corners or abrupt changes in the shape of the solid surface.
FIGURE 11–13 Flow separation in a waterfall. Separation point
Reattachment point
Separated flow region
FIGURE 11–14 Flow separation over a backwardfacing step along a wall.
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570 FLUID MECHANICS
Wake region
FIGURE 11–15 Flow separation during flow over a tennis ball and the wake region. Courtesy NASA and Cislunar Aerospace, Inc.
When a fluid separates from a body, it forms a separated region between the body and the fluid stream. This low-pressure region behind the body where recirculating and backflows occur is called the separated region. The larger the separated region, the larger the pressure drag. The effects of flow separation are felt far downstream in the form of reduced velocity (relative to the upstream velocity). The region of flow trailing the body where the effects of the body on velocity are felt is called the wake (Fig. 11–15). The separated region comes to an end when the two flow streams reattach. Therefore, the separated region is an enclosed volume, whereas the wake keeps growing behind the body until the fluid in the wake region regains its velocity and the velocity profile becomes nearly flat again. Viscous and rotational effects are the most significant in the boundary layer, the separated region, and the wake. The occurrence of separation is not limited to blunt bodies. Complete separation over the entire back surface may also occur on a streamlined body such as an airplane wing at a sufficiently large angle of attack (larger than about 15° for most airfoils), which is the angle the incoming fluid stream makes with the chord (the line that connects the nose and the end) of the wing. Flow separation on the top surface of a wing reduces lift drastically and may cause the airplane to stall. Stalling has been blamed for many airplane accidents and loss of efficiencies in turbomachinery (Fig. 11–16). Note that drag and lift are strongly dependent on the shape of the body, and any effect that causes the shape to change has a profound effect on the drag and lift. For example, snow accumulation and ice formation on airplane wings may change the shape of the wings sufficiently to cause significant loss of lift. This phenomenon has caused many airplanes to lose altitude and crash and many others to abort takeoff. Therefore, it has become a routine safety measure to check for ice or snow buildup on critical components of airplanes before takeoff in bad weather. This is especially important for airplanes that have waited a long time on the runway before takeoff because of heavy traffic. An important consequence of flow separation is the formation and shedding of circulating fluid chunks, called vortices, in the wake region. The
FIGURE 11–16 At large angles of attack (usually larger than 15°), flow may separate completely from the top surface of an airfoil, reducing lift drastically and causing the airfoil to stall. From G. M. Homsy, et al. (2000).
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571 CHAPTER 11
periodic generation of these vortices downstream is referred to as vortex shedding. This phenomenon usually occurs during normal flow over long cylinders or spheres for Re ! 90. The vibrations generated by vortices near the body may cause the body to resonate to dangerous levels if the frequency of the vortices is close to the natural frequency of the body—a situation that must be avoided in the design of equipment that is subjected to high-velocity fluid flow such as the wings of airplanes and suspended bridges subjected to steady high winds.
11–4
■
DRAG COEFFICIENTS OF COMMON GEOMETRIES
The concept of drag has important consequences in daily life, and the drag behavior of various natural and human-made bodies is characterized by their drag coefficients measured under typical operating conditions. Although drag is caused by two different effects (friction and pressure), it is usually difficult to determine them separately. Besides, in most cases, we are interested in the total drag rather than the individual drag components, and thus usually the total drag coefficient is reported. The determination of drag coefficients has been the topic of numerous studies (mostly experimental), and there is a huge amount of drag coefficient data in the literature for just about any geometry of practical interest. The drag coefficient, in general, depends on the Reynolds number, especially for Reynolds numbers below about 104. At higher Reynolds numbers, the drag coefficients for most geometries remain essentially constant (Fig. 11–17). This is due to the flow at high Reynolds numbers becoming fully turbulent. However, this is not the case for rounded bodies such as circular cylinders and spheres, as we discuss later in this section. The reported drag coefficients are usually applicable only to flows at high Reynolds numbers. The drag coefficient exhibits different behavior in the low (creeping), moderate (laminar), and high (turbulent) regions of the Reynolds number. The inertia effects are negligible in low Reynolds number flows (Re " 1), called creeping flows, and the fluid wraps around the body smoothly. The drag coefficient in this case is inversely proportional to the Reynolds number, and for a sphere it is determined to be Sphere:
CD #
24 Re
(Re $ 1)
(11–11)
Then the drag force acting on a spherical object at low Reynolds numbers becomes FD # C D A
rV 2 24 rV 2 24 pD2 rV 2 # A # # 3pmVD 2 Re 2 rVD/m 4 2
(11–12)
which is known as Stokes law, after British mathematician and physicist G. G. Stokes (1819–1903). This relation shows that at very low Reynolds numbers, the drag force acting on spherical objects is proportional to the diameter, the velocity, and the viscosity of the fluid. This relation is often applicable to dust particles in the air and suspended solid particles in water.
CD
Disk
2.0
V
1.5 1.0 0.5
0 101
102
103
104
105
106
FIGURE 11–17 The drag coefficients for most geometries (but not all) remain essentially constant at Reynolds numbers above about 104.
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572 FLUID MECHANICS Sphere
Hemisphere
D
V
CD = 24/Re Circular disk (normal to flow)
FIGURE 11–18 Drag coefficients CD at low velocities (Re ) 1 where Re # VD/n and A # pD2/4).
A hemisphere at two different orientations for Re > 104
V
V
D
V
CD = 22.2/Re Circular disk (parallel to flow)
D
CD = 20.4/Re
V D CD = 13.6/Re
The drag coefficients for low Reynolds number flows past some other geometries are given in Fig. 11–18. Note that at low Reynolds numbers, the shape of the body does not have a major influence on the drag coefficient. The drag coefficients for various two- and three-dimensional bodies are given in Tables 11–1 and 11–2 for large Reynolds numbers. We can make several observations from these tables about the drag coefficient at high Reynolds numbers. First of all, the orientation of the body relative to the direction of flow has a major influence on the drag coefficient. For example, the drag coefficient for flow over a hemisphere is 0.4 when the spherical side faces the flow, but it increases threefold to 1.2 when the flat side faces the flow (Fig. 11–19). This shows that the rounded nose of a bullet serves another purpose in addition to piercing: reducing drag and thus increasing the range of the gun. For blunt bodies with sharp corners, such as flow over a rectangular block or a flat plate normal to flow, separation occurs at the edges of the front and back surfaces, with no significant change in the character of flow. Therefore, the drag coefficient of such bodies is nearly independent of the Reynolds number. Note that the drag coefficient of a long rectangular rod can be reduced almost by half from 2.2 to 1.2 by rounding the corners.
CD = 0.4
Biological Systems and Drag V
CD = 1.2
FIGURE 11–19 The drag coefficient of a body may change drastically by changing the body’s orientation (and thus shape) relative to the direction of flow.
The concept of drag also has important consequences for biological systems. For example, the bodies of fish, especially the ones that swim fast for long distances (such as dolphins), are highly streamlined to minimize drag (the drag coefficient of dolphins based on the wetted skin area is about 0.0035, comparable to the value for a flat plate in turbulent flow). So it is no surprise that we build submarines that mimic large fish. Tropical fish with fascinating beauty and elegance, on the other hand, swim gracefully short distances only. Obviously grace, not high speed and drag, was the primary consideration in their design. Birds teach us a lesson on drag reduction by extending their beak forward and folding their feet backward during flight
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573 CHAPTER 11
TA B L E 1 1 – 1 Drag coefficients CD of various two-dimensional bodies for Re * 104 based on the frontal area A # bD, where b is the length in direction normal to the page (for use in the drag force relation FD # CD ArV 2/2 where V is the upstream velocity) Square rod
Rectangular rod L
V
V
D
r D
Sharp corners: CD = 2.2
V
Round corners (r/D = 0.2): CD = 1.2
D
L V
Circular rod (cylinder) V
D
V
D
CD
0.0* 0.1 0.5 1.0 2.0 3.0
1.9 1.9 2.5 2.2 1.7 1.3
* Corresponds to thin plate
Round front edge:
L/D
CD
0.5 1.0 2.0 4.0
1.2 0.9 0.7 0.7
Elliptical rod Laminar: CD = 1.2 Turbulent: CD = 0.3
Equilateral triangular rod V
Sharp corners:
L/D
CD
L V
D
Semicircular shell
D
CD = 1.5
D
CD = 2.0
V
V
L/D
Laminar
Turbulent
2 4 8
0.60 0.35 0.25
0.20 0.15 0.10
Semicircular rod
D
CD = 2.3
D
CD = 1.2
(Fig. 11–20). Airplanes, which look somewhat like big birds, retract their wheels after takeoff in order to reduce drag and thus fuel consumption. The flexible structure of plants enables them to reduce drag at high winds by changing their shapes. Large flat leaves, for example, curl into a low-drag conical shape at high wind speeds, while tree branches cluster to reduce drag. Flexible trunks bend under the influence of the wind to reduce drag, and the bending moment is lowered by reducing frontal area.
V
V
D
CD = 1.2
D
CD = 1.7
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574 FLUID MECHANICS
TA B L E 1 1 – 2 Representative drag coefficients CD for various three-dimensional bodies for Re * 104 based on the frontal area (for use in the drag force relation FD # CD ArV 2/2 where V is the upstream velocity) Cube, A # D 2
Thin circular disk, A # pD 2/4 V
V
D
Laminar: CD = 0.5 Turbulent: CD = 0.2
D
L
CD
V
D
CD = 0.4 V
V
D
CD = 1.2
L
L/D
Laminar
Turbulent
0.75 1 2 4 8
0.5 0.5 0.3 0.3 0.2
0.2 0.2 0.1 0.1 0.1
Short cylinder, horizontal, A # pD 2/4
Short cylinder, vertical, A # LD D
D
CD = 0.5
D
Ellipsoid, A # pD 2/4
Hemisphere, A # pD 2/4 V
V
CD = 1.1
D
CD = 1.05
Sphere, A # pD 2/4 V
Cone (for u # 30°), A # pD 2/4
L/D
CD
1 2 5 10 40 !
0.6 0.7 0.8 0.9 1.0 1.2
V
D L
L/D
CD
0.5 1 2 4 8
1.1 0.9 0.9 0.9 1.0
Values are for laminar flow
Feet folded back
Beak extended forward
FIGURE 11–20 Birds teach us a lesson on drag reduction by extending their beak forward and folding their feet backward during flight.
If you watch the Olympic games, you have probably observed many instances of conscious effort by the competitors to reduce drag. Some examples: During 100-m running, the runners hold their fingers together and straight and move their hands parallel to the direction of motion to reduce the drag on their hands. Swimmers with long hair cover their head with a tight and smooth cover to reduce head drag. They also wear well-fitting one-piece swimming suits. Horse and bicycle riders lean forward as much as they can to reduce drag (by reducing both the drag coefficient and frontal area). Speed skiers do the same thing. Fairings are commonly used in motorcycles to reduce drag.
Drag Coefficients of Vehicles
The term drag coefficient is commonly used in various areas of daily life. Car manufacturers try to attract consumers by pointing out the low drag coefficients of their cars (Fig. 11–21). The drag coefficients of vehicles
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575 CHAPTER 11
TABLE 11–2 (Continued) Streamlined body, A ! pD 2/4 V
D
Parachute, A ! pD 2/4
Tree, A ! frontal area
D
CD = 0.04
V
V, m/s
CD
10 20 30
0.4–1.2 0.3–1.0 0.2–0.7
CD = 1.3
Person (average)
Bikes CD = 0.5
Standing: CDA ! 9 ft2 ! 0.84 m2 Sitting: CDA ! 6 ft2 ! 0.56 m2
Semitrailer, A ! frontal area
CD = 0.9
Upright: A = 5.5 ft2 = 0.51 m2 CD = 1.1
Drafting: A = 3.9 ft2 = 0.36 m2 CD = 0.50
Racing: A = 3.9 ft2 = 0.36 m2 CD = 0.9
With fairing: A = 5.0 ft2 = 0.46 m2 CD = 0.12
Automotive, A ! frontal area
High-rise buildings, A ! frontal area Minivan: CD = 0.4
CEN_4
Passenger car: CD = 0.3 CD = 1.4
range from about 1.0 for large semitrailers to 0.4 for minivans and 0.3 for passenger cars. In general, the more blunt the vehicle, the higher the drag coefficient. Installing a fairing reduces the drag coefficient of tractor-trailer rigs by about 20 percent by making the frontal surface more streamlined. As a rule of thumb, the percentage of fuel savings due to reduced drag is about half the percentage of drag reduction.
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FIGURE 11–21 Streamlines around an aerodynamically designed modern car closely resemble the streamlines around the car in the ideal potential flow (assumes negligible friction), except near the rear end, resulting in a low drag coefficient. From G. M. Homsy, et al. (2000).
FIGURE 11–22 This sleek-looking 2005 model of the Toyota Prius has a drag coefficient of 0.26—one of the lowest for a passenger car.
When the effect of the road on air motion is disregarded, the ideal shape of a vehicle is the basic teardrop, with a drag coefficient of about 0.1 for the turbulent flow case. But this shape needs to be modified to accommodate several necessary external components such as wheels, mirrors, axles, and door handles. Also, the vehicle must be high enough for comfort and there must be a minimum clearance from the road. Further, a vehicle cannot be too long to fit in garages and parking spaces. Controlling the material and manufacturing costs requires minimizing or eliminating any “dead” volume that cannot be utilized. The result is a shape that resembles more a box than a teardrop, and this was the shape of early cars with a drag coefficient of about 0.8 in the 1920s. This wasn’t a problem in those days since the velocities were low and drag was not a major design consideration. The average drag coefficients of cars dropped to about 0.70 in the 1940s, to 0.55 in the 1970s, to 0.45 in the 1980s, and to 0.30 in the 1990s as a result of improved manufacturing techniques for metal forming and paying more attention to the shape of the car and streamlining (Fig. 11–22). The drag coefficient for well-built racing cars is about 0.2, but this is achieved after making the comfort of drivers a secondary consideration. Noting that the theoretical lower limit of CD is about 0.1 and the value for racing cars is 0.2, it appears that there is only a little room for further improvement in the drag coefficient of passenger cars from the current value of about 0.3. For trucks and buses, the drag coefficient can be reduced further by optimizing the front and rear contours (by rounding, for example) to the extent it is practical while keeping the overall length of the vehicle the same. When traveling as a group, a sneaky way of reducing drag is drafting, a phenomenon well known by bicycle riders and car racers. It involves approaching a moving body from behind and being drafted into the lowpressure region in the rear of the body. The drag coefficient of a racing bicyclist, for example, can be reduced from 0.9 to 0.5 by drafting, as shown in Table 11–2 (Fig. 11–23). We also can help reduce the overall drag of a vehicle and thus fuel consumption by being more conscientious drivers. For example, drag force is proportional to the square of velocity. Therefore, driving over the speed limit on the highways not only increases the chances of getting speeding tickets or getting into an accident, but it also increases the amount of fuel consumption per mile. Therefore, driving at moderate speeds is safe and economical. Also, anything that extends from the car, even an arm, increases
Courtesy Toyota.
FIGURE 11–23 The drag coefficients of bodies following other moving bodies closely can be reduced considerably due to drafting (i.e., falling into the low pressure region created by the body in front).
CD = 0.5
Low pressure
CD = 0.9
High pressure
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577 CHAPTER 11
the drag coefficient. Driving with the windows rolled down also increases the drag and fuel consumption. At highway speeds, a driver can save fuel in hot weather by running the air conditioner instead of driving with the windows rolled down. Usually the turbulence and additional drag generated by open windows consume more fuel than does the air conditioner.
Superposition
The shapes of many bodies encountered in practice are not simple. But such bodies can be treated conveniently in drag force calculations by considering them to be composed of two or more simple bodies. A satellite dish mounted on a roof with a cylindrical bar, for example, can be considered to be a combination of a hemispherical body and a cylinder. Then the drag coefficient of the body can be determined approximately by using superposition. Such a simplistic approach does not account for the effects of components on each other, and thus the results obtained should be interpreted accordingly.
EXAMPLE 11–2
Effect of Mirror Design on the Fuel Consumption of a Car
As part of the continuing efforts to reduce the drag coefficient and thus to improve the fuel efficiency of cars, the design of side rearview mirrors has changed drastically from a simple circular plate to a streamlined shape. Determine the amount of fuel and money saved per year as a result of replacing a 13-cm-diameter flat mirror by one with a hemispherical back (Fig. 11–24). Assume the car is driven 24,000 km a year at an average speed of 95 km/h. Take the density and price of gasoline to be 0.8 kg/L and $0.60/L, respectively; the heating value of gasoline to be 44,000 kJ/kg; and the overall efficiency of the engine to be 30 percent.
SOLUTION The flat mirror of a car is replaced by one with a hemispherical back. The amount of fuel and money saved per year as a result are to be determined. Assumptions 1 The car is driven 24,000 km a year at an average speed of 95 km/h. 2 The effect of the car body on the flow around the mirror is negligible (no interference). Properties The densities of air and gasoline are taken to be 1.20 kg/m3 and 800 kg/m3, respectively. The heating value of gasoline is given to be 44,000 kJ/kg. The drag coefficients CD are 1.1 for a circular disk and 0.4 for a hemispherical body. Analysis The drag force acting on a body is determined from FD # C D A
rV 2 2
where A is the frontal area of the body, which is A # pD2/4 for both the flat and rounded mirrors. The drag force acting on the flat mirror is
FD # 1.1
p(0.13 m)2 (1.20 kg/m3)(95 km/h)2 1 m/s 2 1N a b a b # 6.10 N 4 2 3.6 km/h 1 kg ' m/s2
Flat mirror 95 km/h
Rounded mirror 95 km/h
D = 13 cm
D = 13 cm
FIGURE 11–24 Schematic for Example 11–2.
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578 FLUID MECHANICS
Noting that work is force times distance, the amount of work done to overcome this drag force and the required energy input for a distance of 24,000 km are
Wdrag # FD & L # (6.10 N)(24,000 km/year) # 146,400 kJ/year E in #
Wdrag h car
#
146,400 kJ/year # 488,000 kJ/year 0.3
Then the amount and costs of the fuel that supplies this much energy are
Amount of fuel #
mfuel E in/HV (488,000 kJ/year)/(44,000 kJ/kg) # # r fuel r fuel 0.8 kg/L
# 13.9 L/year Cost # (Amount of fuel)(Unit cost) # (13.9 L/year)($0.60/L) # $8.32/year That is, the car uses 13.9 L of gasoline at a cost of $8.32 per year to overcome the drag generated by a flat mirror extending out from the side of a car. The drag force and the work done to overcome it are directly proportional to the drag coefficient. Then the percent reduction in the fuel consumption due to replacing the mirror is equal to the percent reduction in the drag coefficient:
Reduction ratio #
C D, flat $ C D, hemisp C D, flat
#
1.1 $ 0.4 # 0.636 1.1
Fuel reduction # (Reduction ratio)(Amount of fuel) # 0.636(13.9 L/year) # 8.84 L/year Cost reduction # (Reduction ratio)(Cost) # 0.636($8.32/year) # $5.29/year Since a typical car has two side rearview mirrors, the driver saves more than $10 per year in gasoline by replacing the flat mirrors with hemispherical ones. Discussion Note from this example that significant reductions in drag and fuel consumption can be achieved by streamlining the shape of various components and the entire car. So it is no surprise that the sharp corners are replaced in late model cars by rounded contours. This also explains why large airplanes retract their wheels after takeoff and small airplanes use contoured fairings around their wheels.
Example 11–2 is indicative of the tremendous amount of effort put in recent years into redesigning various parts of cars such as the window moldings, the door handles, the windshield, and the front and rear ends in order to reduce aerodynamic drag. For a car moving on a level road at constant speed, the power developed by the engine is used to overcome rolling resistance, friction between moving components, aerodynamic drag, and driving the auxiliary equipment. The aerodynamic drag is negligible at low speeds, but becomes significant at speeds above about 30 mi/h. Reduction of the frontal area of the cars (to the dislike of tall drivers) has also contributed greatly to the reduction of drag and fuel consumption.
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11–5
■
PARALLEL FLOW OVER FLAT PLATES
Consider the flow of a fluid over a flat plate, as shown in Fig. 11–25. Surfaces that are slightly contoured such as turbine blades can also be approximated as flat plates with reasonable accuracy. The x-coordinate is measured along the plate surface from the leading edge of the plate in the direction of the flow, and y is measured from the surface in the normal direction. The fluid approaches the plate in the x-direction with a uniform velocity V, which is equivalent to the velocity over the plate away from the surface. For the sake of discussion, we can consider the fluid to consist of adjacent layers piled on top of each other. The velocity of the particles in the first fluid layer adjacent to the plate becomes zero because of the no-slip condition. This motionless layer slows down the particles of the neighboring fluid layer as a result of friction between the particles of these two adjoining fluid layers at different velocities. This fluid layer then slows down the molecules of the next layer, and so on. Thus, the presence of the plate is felt up to some normal distance d from the plate beyond which the free-stream velocity remains virtually unchanged. As a result, the x-component of the fluid velocity, u, varies from 0 at y ! 0 to nearly V at y ! d (Fig. 11–26). The region of the flow above the plate bounded by d in which the effects of the viscous shearing forces caused by fluid viscosity are felt is called the velocity boundary layer. The boundary layer thickness d is typically defined as the distance y from the surface at which u ! 0.99V. The hypothetical line of u ! 0.99V divides the flow over a plate into two regions: the boundary layer region, in which the viscous effects and the velocity changes are significant, and the irrotational flow region, in which the frictional effects are negligible and the velocity remains essentially constant. For parallel flow over a flat plate, the pressure drag is zero, and thus the drag coefficient is equal to the friction drag coefficient, or simply the friction coefficient (Fig. 11–27). That is, Flat plate:
CD ! CD, friction ! Cf
(11–13)
Once the average friction coefficient Cf is available, the drag (or friction) force over the surface can be determined from FD ! Ff ! 12C f ArV 2
Friction force on a flat plate: V
Laminar boundary layer
Transition region
(11–14)
Turbulent boundary layer V
V
y
0
x xcr
Turbulent layer
Boundary layer thickness, d
Overlap layer Buffer layer Viscous sublayer
FIGURE 11–25 The development of the boundary layer for flow over a flat plate, and the different flow regimes. Not to scale.
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580 FLUID MECHANICS Relative velocities of fluid layers V
V
Zero velocity at the surface
0.99V d
FIGURE 11–26 The development of a boundary layer on a surface is due to the no-slip condition and friction.
Flow over a flat plate CD, pressure = 0 CD = CD, friction = Cf FD, pressure = 0 FD = FD, friction = Ff = Cf A
rV 2 2
where A is the surface area of the plate exposed to fluid flow. When both sides of a thin plate are subjected to flow, A becomes the total area of the top and bottom surfaces. Note that the local value of friction coefficient Cf, x , in general, varies with location along the surface. Typical average velocity profiles in laminar and turbulent flow are also given in Fig. 11–25. Note that the velocity profile in turbulent flow is much fuller than that in laminar flow, with a sharp drop near the surface. The turbulent boundary layer can be considered to consist of four regions, characterized by the distance from the wall. The very thin layer next to the wall where viscous effects are dominant is the viscous sublayer. The velocity profile in this layer is very nearly linear, and the flow is streamlined. Next to the viscous sublayer is the buffer layer, in which turbulent effects are becoming significant, but the flow is still dominated by viscous effects. Above the buffer layer is the overlap layer, in which the turbulent effects are much more significant, but still not dominant. Above that is the turbulent layer in which turbulent effects dominate over viscous effects. Note that the turbulent boundary layer profile on a flat plate closely resembles the boundary layer profile in fully developed turbulent pipe flow. The transition from laminar to turbulent flow depends on the surface geometry, surface roughness, upstream velocity, surface temperature, and the type of fluid, among other things, and is best characterized by the Reynolds number. The Reynolds number at a distance x from the leading edge of a flat plate is expressed as Rex !
FIGURE 11–27 For parallel flow over a flat plate, the pressure drag is zero, and thus the drag coefficient is equal to the friction coefficient and the drag force is equal to the friction force.
rVx Vx ! m n
(11–15)
where V is the upstream velocity and x is the characteristic length of the geometry, which, for a flat plate, is the length of the plate in the flow direction. Note that unlike pipe flow, the Reynolds number varies for a flat plate along the flow, reaching ReL ! VL/n at the end of the plate. For any point on a flat plate, the characteristic length is the distance x of the point from the leading edge in the flow direction. For flow over a smooth flat plate, transition from laminar to turbulent begins at about Re ! 1 " 105, but does not become fully turbulent before the Reynolds number reaches much higher values, typically around 3 " 106 (Chap. 10). In engineering analysis, a generally accepted value for the critical Reynolds number is Rex, cr !
rVxcr ! 5 " 105 m
The actual value of the engineering critical Reynolds number for a flat plate may vary somewhat from about 105 to 3 " 106 depending on the surface roughness, the turbulence level, and the variation of pressure along the surface, as discussed in more detail in Chap. 10.
Friction Coefficient
The friction coefficient for laminar flow over a flat plate can be determined theoretically by solving the conservation of mass and momentum equations numerically (Chap. 10). For turbulent flow, however, it must be determined experimentally and expressed by empirical correlations.
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581 CHAPTER 11
The local friction coefficient varies along the surface of the flat plate as a result of the changes in the velocity boundary layer in the flow direction. We are usually interested in the drag force on the entire surface, which can be determined using the average friction coefficient. But sometimes we are also interested in the drag force at a certain location, and in such cases, we need to know the local value of the friction coefficient. With this in mind, we present correlations for both local (identified with the subscript x) and average friction coefficients over a flat plate for laminar, turbulent, and combined laminar and turbulent flow conditions. Once the local values are available, the average friction coefficient for the entire plate can be determined by integration from 1 Cf # L
!
L
C f, x dx
d#
4.91x Re1/2 x
and
0.38x Re1/5 x
and
C f, x #
C f, x #
0.664 , Re1/2 x
Rex ( 5 & 10 5
(11–17)
0.059 , Re1/5 x
5 & 10 5 ) Rex ) 10 7
(11–18)
where x is the distance from the leading edge of the plate and Rex # Vx/n is the Reynolds number at location x. Note that Cf, x is proportional to 1/Rex1/2 and thus to x$1/2 for laminar flow and it is proportional to x$1/5 for turbulent flow. In either case, Cf, x is infinite at the leading edge (x # 0), and therefore Eqs. 11–17 and 11–18 are not valid close to the leading edge. The variation of the boundary layer thickness d and the friction coefficient Cf, x along a flat plate is shown in Fig. 11–28. The local friction coefficients are higher in turbulent flow than they are in laminar flow because of the intense mixing that occurs in the turbulent boundary layer. Note that Cf, x reaches its highest values when the flow becomes fully turbulent, and then decreases by a factor of x$1/5 in the flow direction, as shown in the figure. The average friction coefficient over the entire plate is determined by substituting Eqs. 11–17 and 11–18 into Eq. 11–16 and performing the integrations (Fig. 11–29). We get Laminar:
Cf #
1.33 Re1/2 L
Turbulent:
Cf #
0.074 Re1/5 L
d
V
(11–16)
The corresponding relations for turbulent flow are Turbulent: d #
Cf, x
Laminar Transition
Turbulent
0
Based on analysis, the boundary layer thickness and the local friction coefficient at location x for laminar flow over a flat plate were determined in Chap. 10 to be Laminar:
Cf, x
ReL ( 5 & 10 5 5 & 10 5 ) ReL ) 10 7
(11–19)
x
FIGURE 11–28 The variation of the local friction coefficient for flow over a flat plate. Note that the vertical scale of the boundary layer is greatly exaggerated in this sketch.
1 Cf = –– L 1 = –– L
L
!0 Cf, x L
!0
0.664 = ––––– L
dx
0.664 dx ––––– Re1/2 x L
!0
–1/2 Vx a–––– n b dx
V 0.664 a–––– = ––––– n b L
–1/2
× 0.664 aV b = 2––––––– L nL
1/2 –x––– 1 –– 2
L 0
–1/2
1.328 = ––––– Re1/2 L
(11–20)
The first of these relations gives the average friction coefficient for the entire plate when the flow is laminar over the entire plate. The second relation gives the average friction coefficient for the entire plate only when the flow is turbulent over the entire plate, or when the laminar flow region of the plate is negligibly small relative to the turbulent flow region (that is, xcr (( L where the length of the plate xcr over which the flow is laminar can be determined from Recr # 5 & 105 # Vxcr/n).
FIGURE 11–29 The average friction coefficient over a surface is determined by integrating the local friction coefficient over the entire surface. The values shown here are for a laminar flat plate boundary layer.
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582 FLUID MECHANICS
Relative Roughness, e/L
Friction Coefficient, Cf
0.0* 1 & 10$5 1 & 10$4 1 & 10$3
0.0029 0.0032 0.0049 0.0084
* Smooth surface for Re # 107. Others calculated from Eq. 11–23 for fully rough flow.
FIGURE 11–30 For turbulent flow, surface roughness may cause the friction coefficient to increase severalfold.
In some cases, a flat plate is sufficiently long for the flow to become turbulent, but not long enough to disregard the laminar flow region. In such cases, the average friction coefficient over the entire plate is determined by performing the integration in Eq. 11–16 over two parts: the laminar region 0 ) x ) xcr and the turbulent region xcr ( x ) L as 1 Cf # a L
!
xcr
C f, x, laminar dx %
0
!
L
C f, x, turbulent dxb
(11–21)
xcr
Note that we included the transition region with the turbulent region. Again taking the critical Reynolds number to be Recr # 5 & 105 and performing these integrations after substituting the indicated expressions, the average friction coefficient over the entire plate is determined to be Cf #
0.074 1742 $ ReL Re1/5 L
5 & 10 5 ) ReL ) 10 7
(11–22)
The constants in this relation will be different for different critical Reynolds numbers. Also, the surfaces are assumed to be smooth, and the free stream to be of very low turbulence intensity. For laminar flow, the friction coefficient depends on only the Reynolds number, and the surface roughness has no effect. For turbulent flow, however, surface roughness causes the friction coefficient to increase severalfold, to the point that in the fully rough turbulent regime the friction coefficient is a function of surface roughness alone and is independent of the Reynolds number (Fig. 11–30). This is analogous to flow in pipes. A curve fit of experimental data for the average friction coefficient in this regime is given by Schlichting as Fully rough turbulent regime:
e $2.5 C f # a1.89 $ 1.62 log b L
(11–23)
where e is the surface roughness and L is the length of the plate in the flow direction. In the absence of a better one, this relation can be used for turbulent flow on rough surfaces for Re * 106, especially when e/L * 10$4. Friction coefficients Cf for parallel flow over smooth and rough flat plates are plotted in Fig. 11–31 for both laminar and turbulent flows. Note that Cf increases severalfold with roughness in turbulent flow. Also note that Cf is independent of the Reynolds number in the fully rough region. This chart is the flat-plate analog of the Moody chart for pipe flows. EXAMPLE 11–3
Flow of Hot Oil over a Flat Plate
Engine oil at 40°C flows over a 5-m-long flat plate with a free-stream velocity of 2 m/s (Fig. 11–32). Determine the drag force acting on the plate per unit width.
SOLUTION Engine oil flows over a flat plate. The drag force per unit width of the plate is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The critical Reynolds number is Recr # 5 & 105. Properties The density and kinematic viscosity of engine oil at 40°C are r # 876 kg/m3 and n # 2.485 & 10$4 m2/s.
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583 CHAPTER 11 0.014
200 Fully rough
L
e = 300
0.012
500
0.010
1000 0.008 2000
Cf 0.006
5000 10 4 2 × 10 4
0.004
5 × 10 4
Turbulent smooth
2 × 10 5 10 6
Transition
0.002
FIGURE 11–31 Friction coefficient for parallel flow over smooth and rough flat plates.
Laminar 0 10 5
10 6
10 7
10 8
10 9
From White (2003).
ReL
Analysis plate is
Noting that L # 5 m, the Reynolds number at the end of the
ReL #
(2 m/s)(5 m) VL # # 4.024 & 10 4 n 2.485 & 10 $4 m2/s
which is less than the critical Reynolds number. Thus we have laminar flow over the entire plate, and the average friction coefficient is (Fig. 11–29)
# 1.328 & (4.024 & 10 4)$0.5 # 0.00662 C f # 1.328Re$0.5 L Noting that the pressure drag is zero and thus CD # Cf for parallel flow over a flat plate, the drag force acting on the plate per unit width becomes
(876 kg/m3)(2 m/s)2 rV 2 1N FD # C f A b # 58.0 N # 0.00662(5 & 1 m2) a 2 2 1 kg ' m/s2 The total drag force acting on the entire plate can be determined by multiplying the value just obtained by the width of the plate. Discussion The force per unit width corresponds to the weight of a mass of about 6 kg. Therefore, a person who applies an equal and opposite force to the plate to keep it from moving will feel like he or she is using as much force as is necessary to hold a 6-kg mass from dropping.
11–6
■
FLOW OVER CYLINDERS AND SPHERES
Flow over cylinders and spheres is frequently encountered in practice. For example, the tubes in a shell-and-tube heat exchanger involve both internal flow through the tubes and external flow over the tubes, and both flows must be considered in the analysis of the heat exchanger. Also, many sports such as soccer, tennis, and golf involve flow over spherical balls.
V = 2 m/s Oil
A
L=5m
FIGURE 11–32 Schematic for Example 11–3.
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584 FLUID MECHANICS
The characteristic length for a circular cylinder or sphere is taken to be the external diameter D. Thus, the Reynolds number is defined as Re # VD/n where V is the uniform velocity of the fluid as it approaches the cylinder or sphere. The critical Reynolds number for flow across a circular cylinder or sphere is about Recr " 2 & 105. That is, the boundary layer remains laminar for about Re + 2 & 105 and becomes turbulent for Re , 2 & 105. Cross-flow over a cylinder exhibits complex flow patterns, as shown in Fig. 11–33. The fluid approaching the cylinder branches out and encircles the cylinder, forming a boundary layer that wraps around the cylinder. The fluid particles on the midplane strike the cylinder at the stagnation point, bringing the fluid to a complete stop and thus raising the pressure at that point. The pressure decreases in the flow direction while the fluid velocity increases. At very low upstream velocities (Re + 1), the fluid completely wraps around the cylinder and the two arms of the fluid meet on the rear side of the cylinder in an orderly manner. Thus, the fluid follows the curvature of the cylinder. At higher velocities, the fluid still hugs the cylinder on the frontal side, but it is too fast to remain attached to the surface as it approaches the top (or bottom) of the cylinder. As a result, the boundary layer detaches from the surface, forming a separation region behind the cylinder. Flow in the wake region is characterized by periodic vortex formation and pressures much lower than the stagnation point pressure. The nature of the flow across a cylinder or sphere strongly affects the total drag coefficient CD. Both the friction drag and the pressure drag can be significant. The high pressure in the vicinity of the stagnation point and the low pressure on the opposite side in the wake produce a net force on the body in the direction of flow. The drag force is primarily due to friction drag at low Reynolds numbers (Re ( 10) and to pressure drag at high Reynolds numbers (Re * 5000). Both effects are significant at intermediate Reynolds numbers. The average drag coefficients CD for cross-flow over a smooth single circular cylinder and a sphere are given in Fig. 11–34. The curves exhibit different behaviors in different ranges of Reynolds numbers: • For Re + 1, we have creeping flow (Chap. 10), and the drag coefficient decreases with increasing Reynolds number. For a sphere, it is CD # 24/Re. There is no flow separation in this regime.
FIGURE 11–33 Laminar boundary layer separation with a turbulent wake; flow over a circular cylinder at Re # 2000. Courtesy ONERA, photograph by Werlé.
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585 CHAPTER 11 400 200 100 60 40
CD
20 10 6 4 2
FIGURE 11–34 Average drag coefficient for crossflow over a smooth circular cylinder and a smooth sphere.
Smooth cylinder
1 0.6 0.4 Sphere
0.2 0.1 0.06 10–1
100
101
102
103
104
105
106
Re
• At about Re # 10, separation starts occurring on the rear of the body with vortex shedding starting at about Re " 90. The region of separation increases with increasing Reynolds number up to about Re # 103. At this point, the drag is mostly (about 95 percent) due to pressure drag. The drag coefficient continues to decrease with increasing Reynolds number in this range of 10 ( Re ( 103. (A decrease in the drag coefficient does not necessarily indicate a decrease in drag. The drag force is proportional to the square of the velocity, and the increase in velocity at higher Reynolds numbers usually more than offsets the decrease in the drag coefficient.) • In the moderate range of 103 ( Re ( 105, the drag coefficient remains relatively constant. This behavior is characteristic of blunt bodies. The flow in the boundary layer is laminar in this range, but the flow in the separated region past the cylinder or sphere is highly turbulent with a wide turbulent wake. • There is a sudden drop in the drag coefficient somewhere in the range of 105 ( Re ( 106 (usually, at about 2 & 105). This large reduction in CD is due to the flow in the boundary layer becoming turbulent, which moves the separation point further on the rear of the body, reducing the size of the wake and thus the magnitude of the pressure drag. This is in contrast to streamlined bodies, which experience an increase in the drag coefficient (mostly due to friction drag) when the boundary layer becomes turbulent. Flow separation occurs at about u " 80° (measured from the front stagnation point of a cylinder) when the boundary layer is laminar and at about u " 140° when it is turbulent (Fig. 11–35). The delay of separation in turbulent flow is caused by the rapid fluctuations of the fluid in the transverse direction, which enables the turbulent boundary layer to travel farther along the surface before separation occurs, resulting in a narrower wake and a smaller pressure drag. Keep in mind that turbulent flow has a fuller velocity profile as compared to the laminar case, and thus it requires a stronger adverse pressure gradient to overcome the additional momentum close to the wall. In the range of Reynolds numbers where the flow changes from laminar to turbulent, even the drag force FD decreases as the velocity (and thus the Reynolds number) increases. This results in a sudden decrease in drag of a flying body (sometimes called the drag crisis) and instabilities in flight.
From H. Schlichting, Boundary Layer Theory 7e. Copyright © 1979 The McGraw-Hill Companies, Inc. Used by permission.
(a)
(b)
FIGURE 11–35 Flow visualization of flow over (a) a smooth sphere at Re # 15,000, and (b) a sphere at Re # 30,000 with a trip wire. The delay of boundary layer separation is clearly seen by comparing the two photographs. Courtesy ONERA, photograph by Werlé.
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586 FLUID MECHANICS
Effect of Surface Roughness
We mentioned earlier that surface roughness, in general, increases the drag coefficient in turbulent flow. This is especially the case for streamlined bodies. For blunt bodies such as a circular cylinder or sphere, however, an increase in the surface roughness may actually decrease the drag coefficient, as shown in Fig. 11–36 for a sphere. This is done by tripping the boundary layer into turbulence at a lower Reynolds number, and thus causing the fluid to close in behind the body, narrowing the wake and reducing pressure drag considerably. This results in a much smaller drag coefficient and thus drag force for a rough-surfaced cylinder or sphere in a certain range of Reynolds number compared to a smooth one of identical size at the same velocity. At Re ! 2 " 105, for example, CD ! 0.1 for a rough sphere with e/D ! 0.0015, whereas CD ! 0.5 for a smooth one. Therefore, the drag coefficient in this case is reduced by a factor of 5 by simply roughening the surface. Note, however, that at Re ! 106, CD ! 0.4 for a very rough sphere while CD ! 0.1 for the smooth one. Obviously, roughening the sphere in this case will increase the drag by a factor of 4 (Fig. 11–37). The preceding discussion shows that roughening the surface can be used to great advantage in reducing drag, but it can also backfire on us if we are not careful—specifically, if we do not operate in the right range of the Reynolds number. With this consideration, golf balls are intentionally roughened to induce turbulence at a lower Reynolds number to take advantage of the sharp drop in the drag coefficient at the onset of turbulence in the boundary layer (the typical velocity range of golf balls is 15 to 150 m/s, and the Reynolds number is less than 4 " 105). The critical Reynolds number of dimpled golf balls is about 4 " 104. The occurrence of turbulent flow at this Reynolds number reduces the drag coefficient of a golf ball by about half, as shown in Fig. 11–36. For a given hit, this means a longer distance for the ball. Experienced golfers also give the ball a spin during the hit, which helps the rough ball develop a lift and thus travel higher and farther. A similar argument can be given for a tennis ball. For a table tennis ball, however, the distances are
0.6 e = relative roughness D
0.5
CD =
FD 1 rV 2 p D2 4 2
0.4 Golf ball 0.3
0.2
0.1
FIGURE 11–36 The effect of surface roughness on the drag coefficient of a sphere. From Blevins (1984).
0
e = 1.25 × 10–2 D e = 5 × 10–3 D e = 1.5 × 10–3 D 4 × 104
105
e = 0 (smooth) D
2 × 105
4 × 105
VD Re = n
106
4 × 106
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587 CHAPTER 11
very short, and the balls never reach the speeds in the turbulent range. Therefore, the surfaces of table tennis balls are made smooth. Once the drag coefficient is available, the drag force acting on a body in cross-flow can be determined from Eq. 11–5 where A is the frontal area (A ! LD for a cylinder of length L and A ! pD2/4 for a sphere). It should be kept in mind that free-stream turbulence and disturbances by other bodies in the flow (such as flow over tube bundles) may affect the drag coefficients significantly. EXAMPLE 11–4
Drag Force Acting on a Pipe in a River
A 2.2-cm-outer-diameter pipe is to span across a river at a 30-m-wide section while being completely immersed in water (Fig. 11–38). The average flow velocity of water is 4 m/s and the water temperature is 15°C. Determine the drag force exerted on the pipe by the river.
CD Re 2 " 105 106
The drag coefficient corresponding to this value is, from Fig. 11–34, CD ! 1.0. Also, the frontal area for flow past a cylinder is A ! LD. Then the drag force acting on the pipe becomes
(999.1 kg/m3)(4 m/s)2 rV 2 1N ! 1.0(30 " 0.022 m2) a b 2 2 1 kg $ m/s2 ! 5275 N " 5300 N
FD ! C D A
Discussion Note that this force is equivalent to the weight of a mass over 500 kg. Therefore, the drag force the river exerts on the pipe is equivalent to hanging a total of over 500 kg in mass on the pipe supported at its ends 30 m apart. The necessary precautions should be taken if the pipe cannot support this force. If the river were to flow at a faster speed or if turbulent fluctuations in the river were more significant, the drag force would be even larger. Unsteady forces on the pipe might then be significant.
11–7
■
LIFT
Lift was defined earlier as the component of the net force (due to viscous and pressure forces) that is perpendicular to the flow direction, and the lift coefficient was expressed as CL ! 1
FL
2 rV
2
A
(11–6)
Rough Surface, e/D ! 0.0015
0.5 0.1
0.1 0.4
FIGURE 11–37 Surface roughness may increase or decrease the drag coefficient of a spherical object, depending on the value of the Reynolds number.
SOLUTION A pipe is submerged in a river. The drag force that acts on the pipe is to be determined. Assumptions 1 The outer surface of the pipe is smooth so that Fig. 11–34 can be used to determine the drag coefficient. 2 Water flow in the river is steady. 3 The direction of water flow is normal to the pipe. 4 Turbulence in river flow is not considered. Properties The density and dynamic viscosity of water at 15°C are r ! 999.1 kg/m3 and m ! 1.138 " 10#3 kg/m · s. Analysis Noting that D ! 0.022 m, the Reynolds number is VD rVD (999.1 kg/m3)(4 m/s)(0.022 m) ! ! ! 7.73 " 10 4 Re ! m n 1.138 " 10 #3 kg/m $ s
Smooth Surface
River
Pipe 30 m
FIGURE 11–38 Schematic for Example 11–4.
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588 FLUID MECHANICS Planform area, bc
FL
FD Span, b
a
FIGURE 11–39 Definition of various terms associated with an airfoil.
FL
V
Direction of lift FD
a Direction of wall shear
FIGURE 11–40 For airfoils, the contribution of viscous effects to lift is usually negligible since wall shear is parallel to the surfaces and thus nearly normal to the direction of lift.
Angle of attack
Chord, c
where A in this case is normally the planform area, which is the area that would be seen by a person looking at the body from above in a direction normal to the body, and V is the upstream velocity of the fluid (or, equivalently, the velocity of a flying body in a quiescent fluid). For an airfoil of width (or span) b and chord length c (the length between the leading and trailing edges), the planform area is A # bc. The distance between the two ends of a wing or airfoil is called the wingspan or just the span. For an aircraft, the wingspan is taken to be the total distance between the tips of the two wings, which includes the width of the fuselage between the wings (Fig. 11–39). The average lift per unit planform area FL/A is called the wing loading, which is simply the ratio of the weight of the aircraft to the planform area of the wings (since lift equals the weight during flying at constant altitude). Airplane flight is based on lift, and thus developing a better understanding of lift as well as improving the lift characteristics of bodies have been the focus of numerous studies. Our emphasis in this section is on devices such as airfoils that are specifically designed to generate lift while keeping the drag at a minimum. But it should be kept in mind that some devices such as the spoilers and inverted airfoils on racing cars are designed for the opposite purpose of avoiding lift or even generating negative lift to improve traction and control (some early cars actually “took off” at high speeds as a result of the lift produced, which alerted the engineers to come up with ways to reduce lift in their design). For devices that are intended to generate lift such as airfoils, the contribution of viscous effects to lift is usually negligible since the bodies are streamlined, and wall shear is parallel to the surfaces of such devices and thus nearly normal to the direction of lift (Fig. 11–40). Therefore, lift in practice can be taken to be due entirely to the pressure distribution on the surfaces of the body, and thus the shape of the body has the primary influence on lift. Then the primary consideration in the design of airfoils is minimizing the average pressure at the upper surface while maximizing it at the lower surface. The Bernoulli equation can be used as a guide in identifying the high- and low-pressure regions: Pressure is low at locations where the flow velocity is high, and pressure is high at locations where the flow velocity is low. Also, lift is practically independent of the surface roughness since roughness affects the wall shear, not the pressure. The contribution of shear
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589 CHAPTER 11 Stagnation points
Stagnation points
(a) Irrotational flow past a symmetrical airfoil (zero lift)
(b) Irrotational flow past a nonsymmetrical airfoil (zero lift)
Stagnation points
(c) Actual flow past a nonsymmetrical airfoil (positive lift)
FIGURE 11–41 Irrotational and actual flow past symmetrical and nonsymmetrical two-dimensional airfoils.
to lift is usually only significant for very small (lightweight) bodies that can fly at low velocities (and thus very low Reynolds numbers). Noting that the contribution of viscous effects to lift is negligible, we should be able to determine the lift acting on an airfoil by simply integrating the pressure distribution around the airfoil. The pressure changes in the flow direction along the surface, but it remains essentially constant through the boundary layer in a direction normal to the surface (Chap. 10). Therefore, it seems reasonable to ignore the very thin boundary layer on the airfoil and calculate the pressure distribution around the airfoil from the relatively simple potential flow theory (zero vorticity, irrotational flow) for which net viscous forces are zero for flow past an airfoil. The flow fields obtained from such calculations are sketched in Fig. 11–41 for both symmetrical and nonsymmetrical airfoils by ignoring the thin boundary layer. At zero angle of attack, the lift produced by the symmetrical airfoil is zero, as expected because of symmetry, and the stagnation points are at the leading and trailing edges. For the nonsymmetrical airfoil, which is at a small angle of attack, the front stagnation point has moved down below the leading edge, and the rear stagnation point has moved up to the upper surface close to the trailing edge. To our surprise, the lift produced is calculated again to be zero—a clear contradiction of experimental observations and measurements. Obviously, the theory needs to be modified to bring it in line with the observed phenomenon. The source of inconsistency is the rear stagnation point being at the upper surface instead of the trailing edge. This requires the lower side fluid to make a nearly U-turn and flow around the trailing edge toward the stagnation point while remaining attached to the surface, which is a physical impossibility since the observed phenomenon is the separation of flow at sharp turns (imagine a car attempting to make this turn at high speed). Therefore, the lower side fluid separates smoothly off the trailing edge, and the upper side fluid responds by pushing the rear stagnation point downstream. In fact, the stagnation point at the upper surface moves all the way to the trailing edge. This way the two flow streams from the top and the bottom sides of the airfoil meet at the trailing edge, yielding a smooth flow downstream parallel to the trailing edge. Lift is generated because the flow velocity at the top surface is higher, and thus the pressure on that surface is lower due to the Bernoulli effect.
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590 FLUID MECHANICS Clockwise circulation
Counterclockwise circulation
Starting vortex
FIGURE 11–42 Shortly after a sudden increase in angle of attack, a counterclockwise starting vortex is shed from the airfoil, while clockwise circulation appears around the airfoil, causing lift to be generated. 120 100
NACA 64(1) – 412 airfoil Re = 7 × 10 5 Stall
80 60 CL ––– 40 CD 20 0 –20 –40 –8
–4
0 a degrees
4
8
FIGURE 11–43 The variation of the lift-to-drag ratio with angle of attack for a twodimensional airfoil.
The potential flow theory and the observed phenomenon can be reconciled as follows: Flow starts out as predicted by theory, with no lift, but the lower fluid stream separates at the trailing edge when the velocity reaches a certain value. This forces the separated upper fluid stream to close in at the trailing edge, initiating clockwise circulation around the airfoil. This clockwise circulation increases the velocity of the upper stream while decreasing that of the lower stream, causing lift. A starting vortex of opposite sign (counterclockwise circulation) is then shed downstream (Fig. 11–42), and smooth streamlined flow is established over the airfoil. When the potential flow theory is modified by the addition of an appropriate amount of circulation to move the stagnation point down to the trailing edge, excellent agreement is obtained between theory and experiment for both the flow field and the lift. It is desirable for airfoils to generate the most lift while producing the least drag. Therefore, a measure of performance for airfoils is the lift-todrag ratio, which is equivalent to the ratio of the lift-to-drag coefficients CL/CD. This information is provided by either plotting CL versus CD for different values of the angle of attack (a lift–drag polar) or by plotting the ratio CL/CD versus the angle of attack. The latter is done for a particular airfoil design in Fig. 11–43. Note that the CL/CD ratio increases with the angle of attack until the airfoil stalls, and the value of the lift-to-drag ratio can be of the order of 100. One obvious way to change the lift and drag characteristics of an airfoil is to change the angle of attack. On an airplane, for example, the entire plane is pitched up to increase lift, since the wings are fixed relative to the fuselage. Another approach is to change the shape of the airfoil by the use of movable leading edge and trailing edge flaps, as is commonly done in modern large aircraft (Fig. 11–44). The flaps are used to alter the shape of the wings during takeoff and landing to maximize lift and to enable the aircraft to land or take off at low speeds. The increase in drag during this takeoff and landing is not much of a concern because of the relatively short time periods involved. Once at cruising altitude, the flaps are retracted, and the wing is returned to its “normal” shape with minimal drag coefficient and adequate lift coefficient to minimize fuel consumption while cruising at a constant altitude. Note that even a small lift coefficient can generate a large lift force during normal operation because of the large cruising velocities of aircraft and the proportionality of lift to the square of flow velocity. The effects of flaps on the lift and drag coefficients are shown in Fig. 11–45 for an airfoil. Note that the maximum lift coefficient increases from about 1.5 for the airfoil with no flaps to 3.5 for the double-slotted flap case. But also note that the maximum drag coefficient increases from about 0.06
From Abbott, von Doenhoff, and Stivers (1945).
FIGURE 11–44 The lift and drag characteristics of an airfoil during takeoff and landing can be changed by changing the shape of the airfoil by the use of movable flaps. Photo by Yunus Çengel.
(a) Flaps extended (takeoff)
(b) Flaps retracted (cruising)
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591 CHAPTER 11 3.5
3.48
2.5 2.0
Slotted flap
Double-slotted flap
2.5 2.0 Slotted flap 1.5
1.0
–5
3.0
CL
1.52
1.5
0.5
3.5
Double-slotted flap 2.67
3.0
CL
CL max
FIGURE 11–45 Effect of flaps on the lift and drag coefficients of an airfoil.
1.0 Clean (no flap) 0 5 10 15 Angle of attack, a (deg.)
0.5 20
0
Clean (no flap)
0.05 0.10 0.15 0.20 0.25 0.30 CD
From Abbott and von Doenhoff, for NACA 23012 (1959).
for the airfoil with no flaps to about 0.3 for the double-slotted flap case. This is a fivefold increase in the drag coefficient, and the engines must work much harder to provide the necessary thrust to overcome this drag. The angle of attack of the flaps can be increased to maximize the lift coefficient. Also, the leading and trailing edges extend the chord length, and thus enlarge the wing area A. The Boeing 727 uses a triple-slotted flap at the trailing edge and a slot at the leading edge. The minimum flight velocity can be determined from the requirement that the total weight W of the aircraft be equal to lift and CL # CL, max. That is, W # FL # 12 C L, max rV 2min A
→
Vmin #
2W B rC L, max A
(11–24)
For a given weight, the landing or takeoff speed can be minimized by maximizing the product of the lift coefficient and the wing area, CL, max A. One way of doing that is to use flaps, as already discussed. Another way is to control the boundary layer, which can be accomplished simply by leaving flow sections (slots) between the flaps, as shown in Fig. 11–46. Slots are used to prevent the separation of the boundary layer from the upper surface of the wings and the flaps. This is done by allowing air to move from the high-pressure region under the wing into the low-pressure region at the top surface. Note that the lift coefficient reaches its maximum value CL # CL, max, and thus the flight velocity reaches its minimum, at stall conditions, which is a region of unstable operation and must be avoided. The Federal Aviation Administration (FAA) does not allow operation below 1.2 times the stall speed for safety. Another thing we notice from this equation is that the minimum velocity for takeoff or landing is inversely proportional to the square root of density. Noting that air density decreases with altitude (by about 15 percent at 1500 m), longer runways are required at airports at higher altitudes such as Denver to accommodate higher minimum takeoff and landing velocities. The situation becomes even more critical on hot summer days since the density of air is inversely proportional to temperature.
End Effects of Wing Tips
For airplane wings and other airfoils of finite size, the end effects at the tips become important because of the fluid leakage between the lower and upper
Wing Slot
Flap
FIGURE 11–46 A flapped airfoil with a slot to prevent the separation of the boundary layer from the upper surface and to increase the lift coefficient.
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592 FLUID MECHANICS
FIGURE 11–47 Trailing vortices from a rectangular wing with vortex cores leaving the trailing edge at the tips. Courtesy of The Parabolic Press, Stanford, California. Used with permission.
surfaces. The pressure difference between the lower surface (high-pressure region) and the upper surface (low-pressure region) drives the fluid at the tips upward while the fluid is swept toward the back because of the relative motion between the fluid and the wing. This results in a swirling motion that spirals along the flow, called the tip vortex, at the tips of both wings. Vortices are also formed along the airfoil between the tips of the wings. These distributed vortices collect toward the edges after being shed from the trailing edges of the wings and combine with the tip vortices to form two streaks of powerful trailing vortices along the tips of the wings (Figs. 11–47 and 11–48). Trailing vortices generated by large aircraft continue to exist for a long time for long distances (over 10 km) before they gradually disappear due to viscous dissipation. Such vortices and the accompanying downdraft are strong enough to cause a small aircraft to lose control and flip over if it flies through the wake of a larger aircraft. Therefore, following a large aircraft closely (within 10 km) poses a real danger for smaller aircraft. This issue is the controlling factor that governs the spacing of aircraft at takeoff, which limits the flight capacity at airports. In nature, this effect is used to advantage by birds that migrate in V-formation by utilizing the updraft generated by the bird in front. It has been determined that the birds in a typical flock can fly to their destination in V-formation with one-third less energy. Military jets also occasionally fly in V-formation for the same reason. Tip vortices that interact with the free stream impose forces on the wing tips in all directions, including the flow direction. The component of the force in the flow direction adds to drag and is called induced drag. The total drag of a wing is then the sum of the induced drag (3-D effects) and the drag of the airfoil section. The ratio of the square of the average span of an airfoil to the planform area is called the aspect ratio. For an airfoil with a rectangular planform of chord c and span b, it is expressed as AR #
b2 b2 b # # A bc c
(11–25)
Therefore, the aspect ratio is a measure of how narrow an airfoil is in the flow direction. The lift coefficient of wings, in general, increases while the
FIGURE 11–48 A crop duster flies through smoky air to illustrate the tip vortices produced at the tips of the wing. NASA Langley Research Center.
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593 CHAPTER 11
drag coefficient decreases with increasing aspect ratio. This is because a long narrow wing (large aspect ratio) has a shorter tip length and thus smaller tip losses and smaller induced drag than a short and wide wing of the same planform area. Therefore, bodies with large aspect ratios fly more efficiently, but they are less maneuverable because of their larger moment of inertia (owing to the greater distance from the center). Bodies with smaller aspect ratios maneuver better since the wings are closer to the central part. So it is no surprise that fighter planes (and fighter birds like falcons) have short and wide wings while large commercial planes (and soaring birds like albatrosses) have long and narrow wings. The end effects can be minimized by attaching endplates or winglets at the tips of the wings perpendicular to the top surface. The endplates function by blocking some of the leakage around the wing tips, which results in a considerable reduction in the strength of the tip vortices and the induced drag. Wing tip feathers on birds fan out for the same purpose (Fig. 11–49). The development of efficient (low-drag) airfoils was the subject of intense experimental investigations in the 1930s. These airfoils were standardized by the National Advisory Committee for Aeronautics (NACA, which is now NASA), and extensive lists of data on lift coefficients were reported. The variation of the lift coefficient CL with the angle of attack for two airfoils (NACA 0012 and NACA 2412) is given in Fig. 11–50. We make the following observations from this figure: • The lift coefficient increases almost linearly with the angle of attack a, reaches a maximum at about a # 16°, and then starts to decrease sharply. This decrease of lift with further increase in the angle of attack is called stall, and it is caused by flow separation and the formation of a wide wake region over the top surface of the airfoil. Stall is highly undesirable since it also increases drag. • At zero angle of attack (a # 0°), the lift coefficient is zero for symmetrical airfoils but nonzero for nonsymmetrical ones with greater 2.00 CL
(b) Winglets are used on this sailplane to reduce induced drag.
FIGURE 11–49 Induced drag is reduced by (a) wing tip feathers on bird wings and (b) endplates or other disruptions on airplane wings. (a) © Vol. 44/PhotoDisc. (b) Courtesy Schempp-Hirth. Used by permission.
NACA 2412 section a
1.50
(a) A bearded vulture with its wing feathers fanned out during flight.
V
1.00
0.50
0
NACA 0012 section a
–0.50 –5
0
5
V 10
Angle of attack, a, degrees
15
20
FIGURE 11–50 The variation of the lift coefficient with the angle of attack for a symmetrical and a nonsymmetrical airfoil. From Abbott (1932).
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594 FLUID MECHANICS
curvature at the top surface. Therefore, planes with symmetrical wing sections must fly with their wings at higher angles of attack in order to produce the same lift. • The lift coefficient can be increased by severalfold by adjusting the angle of attack (from 0.25 at a # 0° for the nonsymmetrical airfoil to 1.25 at a # 10°). • The drag coefficient also increases with the angle of attack, often exponentially (Fig. 11–51). Therefore, large angles of attack should be used sparingly for short periods of time for fuel efficiency.
Lift Generated by Spinning
0.020 CD 0.016 0.012 0.008
NACA 23015 section V
0.004 0
0
a
16 20 4 8 12 Angle of attack, a (degrees)
FIGURE 11–51 The variation of the drag coefficient of an airfoil with the angle of attack. From Abbott and von Doenhoff (1959).
You have probably experienced giving a spin to a tennis ball or making a drop shot on a tennis or ping-pong ball by giving a fore spin in order to alter the lift characteristics and cause the ball to produce a more desirable trajectory and bounce of the shot. Golf, soccer, and baseball players also utilize spin in their games. The phenomenon of producing lift by the rotation of a solid body is called the Magnus effect after the German scientist Heinrich Magnus (1802–1870), who was the first to study the lift of rotating bodies, which is illustrated in Fig. 11–52 for the simplified case of irrotational (potential) flow. When the ball is not spinning, the lift is zero because of top–bottom symmetry. But when the cylinder is rotated about its axis, the cylinder drags some fluid around because of the no-slip condition and the flow field reflects the superposition of the spinning and nonspinning flows. The stagnation points shift down, and the flow is no longer symmetric about the horizontal plane that passes through the center of the cylinder. The average pressure on the upper half is less than the average pressure at the lower half because of the Bernoulli effect, and thus there is a net upward force (lift) acting on the cylinder. A similar argument can be given for the lift generated on a spinning ball. The effect of the rate of rotation on the lift and drag coefficients of a smooth sphere is shown in Fig. 11–53. Note that the lift coefficient strongly depends on the rate of rotation, especially at low angular velocities. The effect of the rate of rotation on the drag coefficient is small. Roughness also affects
Stagnation points
FIGURE 11–52 Generation of lift on a rotating circular cylinder for the case of “idealized” potential flow (the actual flow involves flow separation in the wake region).
Stagnation points
Lift
High velocity, low pressure
Low velocity, high pressure (a) Potential flow over a stationary cylinder
(b) Potential flow over a rotating cylinder
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595 CHAPTER 11
the drag and lift coefficients. In a certain range of the Reynolds number, roughness produces the desirable effect of increasing the lift coefficient while decreasing the drag coefficient. Therefore, golf balls with the right amount of roughness travel higher and farther than smooth balls for the same hit.
0.8 CD =
FD 1 rV 2 p D2 2 4
0.6 v
Lift and Drag of a Commercial Airplane
A commercial airplane has a total mass of 70,000 kg and a wing planform area of 150 m2 (Fig. 11–54). The plane has a cruising speed of 558 km/h and a cruising altitude of 12,000 m, where the air density is 0.312 kg/m3. The plane has double-slotted flaps for use during takeoff and landing, but it cruises with all flaps retracted. Assuming the lift and the drag characteristics of the wings can be approximated by NACA 23012 (Fig. 11–45), determine (a) the minimum safe speed for takeoff and landing with and without extending the flaps, (b) the angle of attack to cruise steadily at the cruising altitude, and (c) the power that needs to be supplied to provide enough thrust to overcome wing drag.
V CD , CL
EXAMPLE 11–5
D Smooth sphere
0.4
CL =
FL 1 rV 2 p D2 2 4
0.2 VD Re = n = 6 × 104
SOLUTION The cruising conditions of a passenger plane and its wing characteristics are given. The minimum safe landing and takeoff speeds, the angle of attack during cruising, and the power required are to be determined. Assumptions 1 The drag and lift produced by parts of the plane other than the wings, such as the fuselage drag, are not considered. 2 The wings are assumed to be two-dimensional airfoil sections, and the tip effects of the wings are not considered. 3 The lift and the drag characteristics of the wings can be approximated by NACA 23012 so that Fig. 11–45 is applicable. 4 The average density of air on the ground is 1.20 kg/m3. Properties The densities of air are 1.20 kg/m3 on the ground and 0.312 kg/m3 at cruising altitude. The maximum lift coefficients CL, max of the wings are 3.48 and 1.52 with and without flaps, respectively (Fig. 11–45). Analysis (a) The weight and cruising speed of the airplane are
0
0
1
2
3
4
5
1 vD/V 2
FIGURE 11–53 The variation of lift and drag coefficients of a smooth sphere with the nondimensional rate of rotation for Re # VD/n # 6 & 104. From Goldstein (1938).
1N W # mg # (70,000 kg)(9.81 m/s2)a b # 686,700 N 1 kg ' m/s2 1 m/s b # 155 m/s V # (558 km/h)a 3.6 km/h
The minimum velocities corresponding to the stall conditions without and with flaps, respectively, are obtained from Eq. 11–24,
558 km/h 70,000 kg
1 kg ' m/s2 2(686,700 N) 2W Vmin 1 # # b # 70.9 m/s a 3 2 1N B rC L, max 1A B (1.2 kg/m )(1.52)(150 m ) Vmin 2 #
2W
B rCL, max 2A
#
1 kg ' m/s2 2(686,700 N) b # 46.8 m/s a 1N B (1.2 kg/m3)(3.48)(150 m2)
Then the “safe” minimum velocities to avoid the stall region are obtained by multiplying the values above by 1.2:
Without flaps: With flaps:
Vmin 1, safe # 1.2Vmin 1 # 1.2(70.9 m/s) # 85.1 m/s # 306 km/h Vmin 2, safe # 1.2Vmin 2 # 1.2(46.8 m/s) # 56.2 m/s # 202 km/h
150 m2, double-flapped 12,000 m
FIGURE 11–54 Schematic for Example 11–5.
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596 FLUID MECHANICS
since 1 m/s # 3.6 km/h. Note that the use of flaps allows the plane to take off and land at considerably lower velocities, and thus on a shorter runway. (b) When an aircraft is cruising steadily at a constant altitude, the lift must be equal to the weight of the aircraft, FL # W. Then the lift coefficient is determined to be
CL # 1
FL
2 rV
2
A
1 kg ' m/s2 686,700 N b # 1.22 a 3 2 2 1N 2 (0.312 kg/m )(155 m/s) (150 m )
#1
For the case with no flaps, the angle of attack corresponding to this value of CL is determined from Fig. 11–45 to be a " 10°. (c) When the aircraft is cruising steadily at a constant altitude, the net force acting on the aircraft is zero, and thus thrust provided by the engines must be equal to the drag force. The drag coefficient corresponding to the cruising lift coefficient of 1.22 is determined from Fig. 11–45 to be CD " 0.03 for the case with no flaps. Then the drag force acting on the wings becomes FD # C D A
(0.312 kg/m3)(155 m/s)2 rV 2 1 kN # (0.03)(150 m2) a b 2 2 1000 kg ' m/s2
# 16.9 kN
Noting that power is force times velocity (distance per unit time), the power required to overcome this drag is equal to the thrust times the cruising velocity:
1 kW Power # Thrust & Velocity # FDV # (16.9 kN)(155 m/s)a b 1 kN ' m/s # 2620 kW
Therefore, the engines must supply 2620 kW of power to overcome the drag on the wings during cruising. For a propulsion efficiency of 30 percent (i.e., 30 percent of the energy of the fuel is utilized to propel the aircraft), the plane requires energy input at a rate of 8733 kJ/s. Discussion The power determined is the power to overcome the drag that acts on the wings only and does not include the drag that acts on the remaining parts of the aircraft (the fuselage, the tail, etc.). Therefore, the total power required during cruising will be much greater. Also, it does not consider induced drag, which can be dominant during takeoff when the angle of attack is high (Fig. 11–45 is for a 2-D airfoil, and does not include 3-D effects).
EXAMPLE 11–6 4800 rpm
45 mi/h
Ball m = 0.125 lbm
FIGURE 11–55 Schematic for Example 11–6.
Effect of Spin on a Tennis Ball
A tennis ball with a mass of 0.125 lbm and a diameter of 2.52 in is hit at 45 mi/h with a backspin of 4800 rpm (Fig. 11–55). Determine if the ball will fall or rise under the combined effect of gravity and lift due to spinning shortly after being hit in air at 1 atm and 80°F.
SOLUTION A tennis ball is hit with a backspin. It is to be determined whether the ball will fall or rise after being hit. Assumptions 1 The surfaces of the ball are smooth enough for Fig. 11–53 to be applicable. 2 The ball is hit horizontally so that it starts its motion horizontally.
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Properties The density and kinematic viscosity of air at 1 atm and 80°F are r # 0.07350 lbm/ft3 and n # 1.697 & 10$4 ft2/s. Analysis The ball is hit horizontally, and thus it would normally fall under the effect of gravity without the spin. The backspin generates a lift, and the ball will rise if the lift is greater than the weight of the ball. The lift can be determined from
FL # C L A
rV 2 2
where A is the frontal area of the ball, which is A # pD 2/4. The translational and angular velocities of the ball are
5280 ft 1h ba b # 66 ft/s 1 mi 3600 s
V # (45 mi/h)a
Then,
2p rad 1 min v # (4800 rev/min)a ba b # 502 rad/s 1 rev 60 s vD (502 rad/s)(2.52/12 ft) # # 0.80 rad 2V 2(66 ft/s)
From Fig. 11–53, the lift coefficient corresponding to this value is CL # 0.21. Then the lift force acting on the ball is
FL # (0.21)
p(2.52/12 ft)2 (0.0735 lbm/ft3)(66 ft/s)2 1 lbf b a 4 2 32.2 lbm ' ft/s2
# 0.036 lbf
The weight of the ball is
1 lbf W # mg # (0.125 lbm)(32.2 ft/s2)a b # 0.125 lbf 32.2 lbm ' ft/s2
which is more than the lift. Therefore, the ball will drop under the combined effect of gravity and lift due to spinning with a net force of 0.125 $ 0.036 # 0.089 lbf. Discussion This example shows that the ball can be hit much farther by giving it a backspin. Note that a topspin has the opposite effect (negative lift) and speeds up the drop of the ball to the ground. Also, the Reynolds number for this problem is 8 & 104, which is sufficiently close to the 6 & 104 for which Fig. 11–53 is prepared. Also keep in mind that although some spin may increase the distance traveled by a ball, there is an optimal spin that is a function of launch angle, as most golfers are now more aware. Too much spin decreases distance by introducing more induced drag.
No discussion on lift and drag would be complete without mentioning the contributions of Wilbur (1867–1912) and Orville (1871–1948) Wright. The Wright Brothers are truly the most impressive engineering team of all time. Self-taught, they were well informed of the contemporary theory and practice in aeronautics. They both corresponded with other leaders in the field and published in technical journals. While they cannot be credited with developing the concepts of lift and drag, they used them to achieve the first
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FIGURE 11–56 The Wright Brothers take flight at Kitty Hawk. National Air and Space Museum/Smithsonian Institution.
powered, manned, heavier-than-air, controlled flight (Fig. 11–56). They succeeded, while so many before them failed, because they evaluated and designed parts separately. Before the Wrights, experimenters were building and testing whole airplanes. While intuitively appealing, the approach did not allow the determination of how to make the craft better. When a flight lasts only a moment, you can only guess at the weakness in the design. Thus, a new craft did not necessarily perform any better than its predecessor. Testing was simply one belly flop followed by another. The Wrights changed all that. They studied each part using scale and full-size models in wind tunnels and the field. Well before the first powered flyer was assembled, they knew the area required for their best wing shape to support a plane carrying a man and the engine horsepower required to provide an adequate thrust with their improved impeller. The Wright Brothers not only showed the world how to fly, they showed engineers how to use the equations presented here to design even better aircraft.
SUMMARY In this chapter, we study flow of fluids over immersed bodies with emphasis on the resulting lift and drag forces. A fluid may exert forces and moments on a body in and about various directions. The force a flowing fluid exerts on a body in the flow direction is called drag while that in the direction normal to the flow is called lift. The part of drag that is due directly to wall shear stress tw is called the skin friction drag since it is caused by frictional effects, and the part that is due directly to pressure P is called the pressure drag or form drag because of its strong dependence on the form or shape of the body. The drag coefficient CD and the lift coefficient CL are dimensionless numbers that represent the drag and the lift characteristics of a body and are defined as CD # 1
FD
2 rV
2
A
and
CL # 1
FL
2 rV
2
A
where A is usually the frontal area (the area projected on a plane normal to the direction of flow) of the body. For plates and airfoils, A is taken to be the planform area, which is the area that would be seen by a person looking at the body from directly above. The drag coefficient, in general, depends on the Reynolds number, especially for Reynolds numbers below 104. At higher Reynolds numbers, the drag coefficients for most geometries remain essentially constant. A body is said to be streamlined if a conscious effort is made to align its shape with the anticipated streamlines in the flow in order to reduce drag. Otherwise, a body (such as a building) tends to block the flow and is said to be blunt or bluff. At sufficiently high velocities, the fluid stream detaches itself from the surface of the body. This is called flow separation. When a fluid stream separates from the body, it forms a separated region between the body and the fluid stream. Sep-
aration may also occur on a streamlined body such as an airplane wing at a sufficiently large angle of attack, which is the angle the incoming fluid stream makes with the chord (the line that connects the nose and the end) of the body. Flow separation on the top surface of a wing reduces lift drastically and may cause the airplane to stall. The region of flow above a surface in which the effects of the viscous shearing forces caused by fluid viscosity are felt is called the velocity boundary layer or just the boundary layer. The thickness of the boundary layer, d, is defined as the distance from the surface at which the velocity is 0.99V. The hypothetical line of velocity 0.99V divides the flow over a plate into two regions: the boundary layer region, in which the viscous effects and the velocity changes are significant, and the irrotational outer flow region, in which the frictional effects are negligible and the velocity remains essentially constant. For external flow, the Reynolds number is expressed as ReL #
rVL VL # m n
where V is the upstream velocity and L is the characteristic length of the geometry, which is the length of the plate in the flow direction for a flat plate and the diameter D for a cylinder or sphere. The average friction coefficients over an entire flat plate are Laminar flow:
Cf #
1.33 Re1/2 L
Turbulent flow:
Cf #
0.074 Re1/5 L
ReL ( 5 & 10 5 5 & 10 5 ) ReL ) 10 7
If the flow is approximated as laminar up to the engineering critical number of Recr # 5 & 105, and then turbulent
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beyond, the average friction coefficient over the entire flat plate becomes Cf #
0.074 1742 $ ReL Re1/5 L
5 & 10 5 ) ReL ) 10 7
A curve fit of experimental data for the average friction coefficient in the fully rough turbulent regime is Rough surface:
e $2.5 C f # a1.89 $ 1.62 log b L
where e is the surface roughness and L is the length of the plate in the flow direction. In the absence of a better one, this relation can be used for turbulent flow on rough surfaces for Re * 106, especially when e/L * 10$4. Surface roughness, in general, increases the drag coefficient in turbulent flow. For blunt bodies such as a circular cylinder or sphere, however, an increase in the surface roughness may decrease the drag coefficient. This is done by tripping the flow into turbulence at a lower Reynolds number, and thus causing the fluid to close in behind the body, narrowing the wake and reducing pressure drag considerably. It is desirable for airfoils to generate the most lift while producing the least drag. Therefore, a measure of performance for airfoils is the lift-to-drag ratio, CL/CD.
The minimum safe flight velocity of an aircraft can be determined from Vmin #
2W B rCL, max A
For a given weight, the landing or takeoff speed can be minimized by maximizing the product of the lift coefficient and the wing area, CL, max A. For airplane wings and other airfoils of finite size, the pressure difference between the lower and the upper surfaces drives the fluid at the tips upward. This results in a swirling eddy, called the tip vortex. Tip vortices that interact with the free stream impose forces on the wing tips in all directions, including the flow direction. The component of the force in the flow direction adds to drag and is called induced drag. The total drag of a wing is then the sum of the induced drag (3-D effects) and the drag of the airfoil section. It is observed that lift develops when a cylinder or sphere in flow is rotated at a sufficiently high rate. The phenomenon of producing lift by the rotation of a solid body is called the Magnus effect. Some external flows, complete with flow details including plots of velocity fields, are solved using computational fluid dynamics, and presented in Chap. 15.
REFERENCES AND SUGGESTED READING 1. I. H. Abbott. “The Drag of Two Streamline Bodies as Affected by Protuberances and Appendages,” NACA Report 451, 1932. 2. I. H. Abbott and A. E. von Doenhoff. Theory of Wing Sections, Including a Summary of Airfoil Data. New York: Dover, 1959. 3. I. H. Abbott, A. E. von Doenhoff, and L. S. Stivers. “Summary of Airfoil Data,” NACA Report 824, Langley Field, VA, 1945. 4. J. D. Anderson. Fundamentals of Aerodynamics, 2nd ed. New York: McGraw-Hill, 1991. 5. R. D. Blevins. Applied Fluid Dynamics Handbook. New York: Van Nostrand Reinhold, 1984. 6. S. W. Churchill and M. Bernstein. “A Correlating Equation for Forced Convection from Gases and Liquids to a Circular Cylinder in Cross Flow,” Journal of Heat Transfer 99, pp. 300–306, 1977. 7. C. T. Crowe, J. A. Roberson, and D. F. Elger. Engineering Fluid Mechanics, 7th ed. New York: Wiley, 2001. 8. S. Goldstein. Modern Developments in Fluid Dynamics. London: Oxford Press, 1938. 9. J. Happel. Low Reynolds Number Hydrocarbons. Englewood Cliffs, NJ: Prentice Hall, 1965.
10. S. F. Hoerner. Fluid-Dynamic Drag. [Published by the author.] Library of Congress No. 64, 1966. 11. G. M. Homsy, H. Aref, K. S. Breuer, S. Hochgreb, J. R. Koseff, B. R. Munson, K. G. Powell, C. R. Robertson, S. T. Thoroddsen. Multi-Media Fluid Mechanics (CD). Cambridge University Press, 2000. 12. W. H. Hucho. Aerodynamics of Road Vehicles. London: Butterworth-Heinemann, 1987. 13. B. R. Munson, D. F. Young, and T. Okiishi. Fundamentals of Fluid Mechanics, 4th ed. New York: Wiley, 2002. 14. M. C. Potter and D. C. Wiggert. Mechanics of Fluids, 2nd ed. Upper Saddle River, NJ: Prentice Hall, 1997. 15. C. T. Crowe, J. A. Roberson, and D. F. Elger. Engineering Fluid Mechanics, 7th ed. New York: Wiley, 2001. 16. H. Schlichting. Boundary Layer Theory, 7th ed. New York: McGraw-Hill, 1979. 17. M. Van Dyke. An Album of Fluid Motion. Stanford, CA: The Parabolic Press, 1982. 18. J. Vogel. Life in Moving Fluids, 2nd ed. Boston: Willard Grand Press, 1994. 19. F. M. White. Fluid Mechanics, 5th ed. New York: McGraw-Hill, 2003.
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APPLICATION SPOTLIGHT
■
Drag Reduction
Guest Author: Werner J. A. Dahm, The University of Michigan
FIGURE 11–57 Drag-reducing microactuator arrays on the hull of a submarine. Shown is the system architecture with tiles composed of unit cells containing sensors and actuators.
FIGURE 11–58 Microelectrokinetic actuator array (MEKA-5) with 25,600 individual actuators at 325-mm spacing for fullscale hydronautical drag reduction. Close-up of a single unit cell (top) and partial view of the full array (bottom).
A reduction of just a few percent in the drag that acts on an air vehicle, a naval surface vehicle, or an undersea vehicle can translate into large reductions in fuel weight and operating costs, or increases in vehicle range and payload. One approach to achieve such drag reduction is to actively control naturally occurring streamwise vortices in the viscous sublayer of the turbulent boundary layer at the vehicle surface. The thin viscous sublayer at the base of any turbulent boundary layer is a powerful nonlinear system, capable of amplifying small microactuator-induced perturbations into large reductions in the vehicle drag. Numerous experimental, computational, and theoretical studies have shown that reductions of 15 to 25 percent in the wall shear stress are possible by properly controlling these sublayer structures. The challenge has been to develop large, dense arrays of microactuators that can manipulate these structures to achieve drag reduction on practical aeronautical and hydronautical vehicles (Fig. 11–57). The sublayer structures are typically a few hundred microns, and thus well matched to the scale of microelectromechanical systems (MEMS). Figure 11–58 shows an example of one type of such microscale actuator array based on the electrokinetic principle that is potentially suitable for active sublayer control on real vehicles. Electrokinetic flow provides a way to move small amounts of fluid on very fast time scales in very small devices. The actuators impulsively displace a fixed volume of fluid between the wall and the viscous sublayer in a manner that counteracts the effect of the sublayer vortices. A system architecture based on independent unit cells, appropriate for large arrays of such microactuators, provides greatly reduced control processing requirements within individual unit cells, which consist of a relatively small number of individual sensors and actuators. Fundamental consideration of the scaling principles governing electrokinetic flow, as well as the sublayer structure and dynamics and microfabrication technologies, have been used to develop and produce full-scale electrokinetic microactuator arrays that can meet many of the requirements for active sublayer control of turbulent boundary layers under real-vehicle conditions. Such microelectrokinetic actuator (MEKA) arrays, when fabricated with wall shear stress sensors also based on microelectromechanical systems fabrication, may in the future allow engineers to achieve dramatic reductions in the drag acting on practical aeronautical and hydronautical vehicles. References Diez-Garias, F. J., Dahm, W. J. A., and Paul, P. H., “Microactuator Arrays for Sublayer Control in Turbulent Boundary Layers Using the Electrokinetic Principle,” AIAA Paper No. 2000-0548, AIAA, Washington, DC, 2000. Diez, F. J., and Dahm, W. J. A., “Electrokinetic Microactuator Arrays and System Architecture for Active Sublayer Control of Turbulent Boundary Layers,” AIAA Journal, Vol. 41, pp. 1906–1915, 2003.
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PROBLEMS* Drag, Lift, and Drag Coefficients 11–1C Explain when an external flow is two-dimensional, three-dimensional, and axisymmetric. What type of flow is the flow of air over a car? 11–2C What is the difference between the upstream velocity and the free-stream velocity? For what types of flow are these two velocities equal to each other?
11–15C Fairings are attached to the front and back of a cylindrical body to make it look more streamlined. What is the effect of this modification on the (a) friction drag, (b) pressure drag, and (c) total drag? Assume the Reynolds number is high enough so that the flow is turbulent for both cases. V
Fairings
11–3C What is the difference between streamlined and blunt bodies? Is a tennis ball a streamlined or blunt body? 11–4C Name some applications in which a large drag is desired. 11–5C What is drag? What causes it? Why do we usually try to minimize it? 11–6C What is lift? What causes it? Does wall shear contribute to the lift? 11–7C During flow over a given body, the drag force, the upstream velocity, and the fluid density are measured. Explain how you would determine the drag coefficient. What area would you use in the calculations? 11–8C During flow over a given slender body such as a wing, the lift force, the upstream velocity, and the fluid density are measured. Explain how you would determine the lift coefficient. What area would you use in the calculations? 11–9C Define the frontal area of a body subjected to external flow. When is it appropriate to use the frontal area in drag and lift calculations? 11–10C Define the planform area of a body subjected to external flow. When is it appropriate to use the planform area in drag and lift calculations? 11–11C What is terminal velocity? How is it determined? 11–12C What is the difference between skin friction drag and pressure drag? Which is usually more significant for slender bodies such as airfoils? 11–13C What is the effect of surface roughness on the friction drag coefficient in laminar and turbulent flows? 11–14C In general, how does the drag coefficient vary with the Reynolds number at (a) low and moderate Reynolds numbers and (b) at high Reynolds numbers (Re * 104)?
* Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
Cylinder
FIGURE P11–15C 11–16C What is the effect of streamlining on (a) friction drag and (b) pressure drag? Does the total drag acting on a body necessarily decrease as a result of streamlining? Explain. 11–17C What is flow separation? What causes it? What is the effect of flow separation on the drag coefficient? 11–18C What is drafting? How does it affect the drag coefficient of the drafted body? 11–19C Which car is more likely to be more fuel-efficient: the one with sharp corners or the one that is contoured to resemble an ellipse? Why? 11–20C Which bicyclist is more likely to go faster: the one who keeps his head and his body in the most upright position or the one who leans down and brings his body closer to his knees? Why? 11–21 The drag coefficient of a car at the design conditions of 1 atm, 25°C, and 90 km/h is to be determined experimentally in a large wind tunnel in a full-scale test. The height and width of the car are 1.40 m and 1.65 m, respectively. If the horizontal force acting on the car is measured to be 300 N, determine the total drag coefficient of this car. Answer: 0.35 11–22 A car is moving at a constant velocity of 80 km/h. Determine the upstream velocity to be used in fluid flow analysis if (a) the air is calm, (b) wind is blowing against the direction of motion of the car at 30 km/h, and (c) wind is blowing in the same direction of motion of the car at 50 km/h. 11–23 The resultant of the pressure and wall shear forces acting on a body is measured to be 700 N, making 35° with V
FR = 700 N 35°
FIGURE P11–23
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the direction of flow. Determine the drag and the lift forces acting on the body. 11–24 During a high Reynolds number experiment, the total drag force acting on a spherical body of diameter D # 12 cm subjected to airflow at 1 atm and 5°C is measured to be 5.2 N. The pressure drag acting on the body is calculated by integrating the pressure distribution (measured by the use of pressure sensors throughout the surface) to be 4.9 N. Determine the friction drag coefficient of the sphere.
from all four sides. Determine the increase in the annual fuel cost of this taxicab due to this sign. Assume the taxicab is driven 60,000 km a year at an average speed of 50 km/h and the overall efficiency of the engine is 28 percent. Take the density, unit price, and heating value of gasoline to be 0.75 kg/L, $0.50/L, and 42,000 kJ/kg, respectively, and the density of air to be 1.25 kg/m3. Pa’s Pizza
Answer: 0.0115
11–25E To reduce the drag coefficient and thus to improve the fuel efficiency, the frontal area of a car is to be reduced. Determine the amount of fuel and money saved per year as a result of reducing the frontal area from 18 to 15 ft2. Assume the car is driven 12,000 mi a year at an average speed of 55 mi/h. Take the density and price of gasoline to be 50 lbm/ft3 and $2.20/gal, respectively; the density of air to be 0.075 lbm/ft3, the heating value of gasoline to be 20,000 Btu/lbm; and the overall efficiency of the engine to be 32 percent. 11–26E
Reconsider Prob. 11–25E. Using EES (or other) software, investigate the effect of frontal area on the annual fuel consumption of the car. Let the frontal area vary from 10 to 30 ft2 in increments of 2 ft2. Tabulate and plot the results.
TAXI FIGURE P11–29 11–30 It is proposed to meet the water needs of a recreational vehicle (RV) by installing a 2-m-long, 0.5-m-diameter cylindrical tank on top of the vehicle. Determine the additional power requirement of the RV at a speed of 95 km/h when the tank is installed such that its circular surfaces face (a) the front and back and (b) the sides of the RV. Assume atmospheric conditions are 87 kPa and 20°C. Answers: (a) 1.67 kW, (b) 7.55 kW 2m
0.5 m
11–27
A circular sign has a diameter of 50 cm and is subjected to normal winds up to 150 km/h at 10°C and 100 kPa. Determine the drag force acting on the sign. Also determine the bending moment at the bottom of its pole whose height from the ground to the bottom of the sign is 1.5 m. Disregard the drag on the pole. 150 km/h
SIGN
1.5 m
FIGURE P11–27 11–28E Wind loading is a primary consideration in the design of the supporting mechanisms of billboards, as evidenced by many billboards being knocked down during high winds. Determine the wind force acting on an 8-ft-high, 20-ft-wide billboard due to 90-mi/h winds in the normal direction when the atmospheric conditions are 14.3 psia and 40°F. Answer: 6684 lbf 11–29 Advertisement signs are commonly carried by taxicabs for additional income, but they also increase the fuel cost. Consider a sign that consists of a 0.30-m-high, 0.9-m-wide, and 0.9-m-long rectangular block mounted on top of a taxicab such that the sign has a frontal area of 0.3 m by 0.9 m
FIGURE P11–30 11–31E At highway speeds, about half of the power generated by the car’s engine is used to overcome aerodynamic drag, and thus the fuel consumption is nearly proportional to the drag force on a level road. Determine the percentage increase in fuel consumption of a car per unit time when a person who normally drives at 55 mi/h now starts driving at 75 mi/h. 11–32 A 4-mm-diameter plastic sphere whose density is 1150 kg/m3 is dropped into water at 20°C. Determine the terminal velocity of the sphere in water. 11–33 During major windstorms, high vehicles such as RVs and semis may be thrown off the road and boxcars off their tracks, especially when they are empty and in open areas. Consider a 5000-kg semi that is 8 m long, 2 m high, and 2 m wide. The distance between the bottom of the truck and the road is 0.75 m. Now the truck is exposed to winds from its side surface. Determine the wind velocity that will tip the truck over to its side. Take the air density to be 1.1 kg/m3 and assume the weight to be uniformly distributed.
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2m
2m
0.75 m
FIGURE P11–33 11–34 An 80-kg bicyclist is riding her 15-kg bicycle downhill on a road with a slope of 12° without pedaling or braking. The bicyclist has a frontal area of 0.45 m2 and a drag coefficient of 1.1 in the upright position, and a frontal area of 0.4 m2 and a drag coefficient of 0.9 in the racing position. Disregarding the rolling resistance and friction at the bearings, determine the terminal velocity of the bicyclist for both positions. Take the air density to be 1.25 kg/m3. Answers: 90 km/h, 106 km/h
11–35 A wind turbine with two or four hollow hemispherical cups connected to a pivot is commonly used to measure wind speed. Consider a wind turbine with two 8-cm-diameter cups with a center-to-center distance of 25 cm, as shown in Fig. P11–35. The pivot is stuck as a result of some malfunction, and the cups stop rotating. For a wind speed of 15 m/s and air density of 1.25 kg/m3, determine the maximum torque this turbine applies on the pivot.
11–38 During steady motion of a vehicle on a level road, the power delivered to the wheels is used to overcome aerodynamic drag and rolling resistance (the product of the rolling resistance coefficient and the weight of the vehicle), assuming the friction at the bearings of the wheels is negligible. Consider a car that has a total mass of 950 kg, a drag coefficient of 0.32, a frontal area of 1.8 m2, and a rolling resistance coefficient of 0.04. The maximum power the engine can deliver to the wheels is 80 kW. Determine (a) the speed at which the rolling resistance is equal to the aerodynamic drag force and (b) the maximum speed of this car. Take the air density to be 1.20 kg/m3. 11–39
Reconsider Prob. 11–38. Using EES (or other) software, investigate the effect of car speed on the required power to overcome (a) rolling resistance, (b) the aerodynamic drag, and (c) their combined effect. Let the car speed vary from 0 to 150 km/h in increments of 15 km/h. Tabulate and plot the results. 11–40 A submarine can be treated as an ellipsoid with a diameter of 5 m and a length of 25 m. Determine the power required for this submarine to cruise horizontally and steadily at 40 km/h in seawater whose density is 1025 kg/m3. Also determine the power required to tow this submarine in air whose density is 1.30 kg/m3. Assume the flow is turbulent in both cases.
40 km/h
Submarine
25 cm
FIGURE P11–40
FIGURE P11–35 11–36
Reconsider Prob. 11–35. Using EES (or other) software, investigate the effect of wind speed on the torque applied on the pivot. Let the wind speed vary from 0 to 50 m/s in increments of 5 m/s. Tabulate and plot the results. 11–37E A 5-ft-diameter spherical tank completely submerged in freshwater is being towed by a ship at 12 ft/s. Assuming turbulent flow, determine the required towing power.
11–41 An 0.80-m-diameter, 1.2-m-high garbage can is found in the morning tipped over due to high winds during the night. Assuming the average density of the garbage inside to be 150 kg/m3 and taking the air density to be 1.25 kg/m3, estimate the wind velocity during the night when the can was tipped over. Take the drag coefficient of the can to be 0.7. Answer: 186 km/h
11–42E The drag coefficient of a vehicle increases when its windows are rolled down or its sunroof is opened. A sports car has a frontal area of 18 ft2 and a drag coefficient of 0.32 when the windows and sunroof are closed. The drag coefficient increases to 0.41 when the sunroof is open. Determine the additional power consumption of the car when the sunroof is opened at (a) 35 mi/h and (b) 70 mi/h. Take the density of air to be 0.075 lbm/ft3.
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house. Assuming the wall surfaces to be smooth, determine the friction drag acting on the wall. What would your answer be if the wind velocity has doubled? How realistic is it to treat the flow over side wall surfaces as flow over a flat plate?
Sunroof closed
Answers: 16 N, 58 N
11–50E
CD # 0.32
Sunroof open
CD # 0.41
Air at 70°F flows over a 10-ft-long flat plate at 25 ft/s. Determine the local friction coefficient at intervals of 1 ft and plot the results against the distance from the leading edge. 11–51 The forming section of a plastics plant puts out a continuous sheet of plastic that is 1.2 m wide and 2 mm thick at a rate of 15 m/min. The sheet is subjected to airflow at a velocity of 3 m/s on both sides along its surfaces normal to the direction of motion of the sheet. The width of the air cooling section is such that a fixed point on the plastic sheet passes through that section in 2 s. Using properties of air at 1 atm and 60°C, determine the drag force the air exerts on the plastic sheet in the direction of airflow.
FIGURE P11–42E Air 3 m/s
Flow over Flat Plates 11–43C What fluid property is responsible for the development of the velocity boundary layer? What is the effect of the velocity on the thickness of the boundary layer?
Plastic sheet
11–44C What does the friction coefficient represent in flow over a flat plate? How is it related to the drag force acting on the plate? 11–45C Consider laminar flow over a flat plate. How does the local friction coefficient change with position? 11–46C How is the average friction coefficient determined in flow over a flat plate? 11–47E Light oil at 75°F flows over a 15-ft-long flat plate with a free-stream velocity of 6 ft/s. Determine the total drag force per unit width of the plate. 11–48 The local atmospheric pressure in Denver, Colorado (elevation 1610 m) is 83.4 kPa. Air at this pressure and at 25°C flows with a velocity of 6 m/s over a 2.5-m & 8-m flat plate. Determine the drag force acting on the top surface of the plate if the air flows parallel to the (a) 8-m-long side and (b) the 2.5-m-long side.
15 m/min
FIGURE P11–51 11–52 The top surface of the passenger car of a train moving at a velocity of 70 km/h is 3.2 m wide and 8 m long. If the outdoor air is at 1 atm and 25°C, determine the drag force acting on the top surface of the car. Air 25°C
70 km/h
11–49 During a winter day, wind at 55 km/h, 5°C, and 1 atm is blowing parallel to a 4-m-high and 10-m-long wall of a
FIGURE P11–52
Air 5°C 55 km/h
4m
10 m
FIGURE P11–49
11–53 The weight of a thin flat plate 50 cm & 50 cm in size is balanced by a counterweight that has a mass of 2 kg, as shown in Fig. P11–53. Now a fan is turned on, and air at 1 atm and 25°C flows downward over both surfaces of the plate with a free-stream velocity of 10 m/s. Determine the mass of the counterweight that needs to be added in order to balance the plate in this case.
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turbulence supposed to increase the drag coefficient instead of decreasing it?
Air 25°C, 10 m/s
11–60C In flow over blunt bodies such as a cylinder, how does the pressure drag differ from the friction drag? 11–61C Why is flow separation in flow over cylinders delayed in turbulent flow? Plate
50 cm
50 cm
FIGURE P11–53 11–54 Consider laminar flow of a fluid over a flat plate. Now the free-stream velocity of the fluid is doubled. Determine the change in the drag force on the plate. Assume the flow to remain laminar. Answer: A 2.83-fold increase 11–55E Consider a refrigeration truck traveling at 65 mi/h at a location where the air temperature is at 1 atm and 80°F. The refrigerated compartment of the truck can be considered to be a 9-ft-wide, 8-ft-high, and 20-ft-long rectangular box. Assuming the airflow over the entire outer surface to be turbulent and attached (no flow separation), determine the drag force acting on the top and side surfaces and the power required to overcome this drag.
20 ft 8 ft
11–62E A 1.2-in-outer-diameter pipe is to span across a river at a 105-ft-wide section while being completely immersed in water. The average flow velocity of water is 10 ft/s, and the water temperature is 70°F. Determine the drag force exerted on the pipe by the river. Answer: 1320 lbf 11–63 A long 8-cm-diameter steam pipe passes through some area open to the winds. Determine the drag force acting on the pipe per unit of its length when the air is at 1 atm and 5°C and the wind is blowing across the pipe at a velocity of 50 km/h. 11–64E A person extends his uncovered arms into the windy air outside at 1 atm and 60°F and 20 mi/h in order to feel nature closely. Treating the arm as a 2-ft-long and 3-indiameter cylinder, determine the drag force on both arms. Answer: 1.02 lbf Air 60°F, 20 mi/h
Air, 80°F V # 65 mi/h
Refrigeration truck
FIGURE P11–55E 11–56E
Reconsider Prob. 11–55E. Using EES (or other) software, investigate the effect of truck speed on the total drag force acting on the top and side surfaces, and the power required to overcome it. Let the truck speed vary from 0 to 100 mi/h in increments of 10 mi/h. Tabulate and plot the results. 11–57 Air at 25°C and 1 atm is flowing over a long flat plate with a velocity of 8 m/s. Determine the distance from the leading edge of the plate where the flow becomes turbulent, and the thickness of the boundary layer at that location. 11–58
Repeat Prob. 11–57 for water.
Flow across Cylinders and Spheres 11–59C In flow over cylinders, why does the drag coefficient suddenly drop when the flow becomes turbulent? Isn’t
FIGURE P11–64E 11–65 A 6-mm-diameter electrical transmission line is exposed to windy air. Determine the drag force exerted on a 120-m-long section of the wire during a windy day when the air is at 1 atm and 15°C and the wind is blowing across the transmission line at 40 km/h. 11–66 Consider 0.8-cm-diameter hail that is falling freely in atmospheric air at 1 atm and 5°C. Determine the terminal velocity of the hail. Take the density of hail to be 910 kg/m3. 11–67 A 0.1-mm-diameter dust particle whose density is 2.1 g/cm3 is observed to be suspended in the air at 1 atm and 25°C at a fixed point. Estimate the updraft velocity of air motion at that location. Assume Stokes law to be applicable. Is this a valid assumption? Answer: 0.62 m/s
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11–68 Dust particles of diameter 0.05 mm and density 1.8 g/cm3 are unsettled during high winds and rise to a height of 350 m by the time things calm down. Estimate how long it will take for the dust particles to fall back to the ground in still air at 1 atm and 15°C, and their velocity. Disregard the initial transient period during which the dust particles accelerate to their terminal velocity, and assume Stokes law to be applicable. 11–69
A 2-m-long, 0.2-m-diameter cylindrical pine log (density # 513 kg/m3) is suspended by a crane in the horizontal position. The log is subjected to normal winds of 40 km/h at 5°C and 88 kPa. Disregarding the weight of the cable and its drag, determine the angle u the cable will make with the horizontal and the tension on the cable.
Lift 11–71C Why is the contribution of viscous effects to lift usually negligible for airfoils? 11–72C Air is flowing past a symmetrical airfoil at zero angle of attack. Will the (a) lift and (b) drag acting on the airfoil be zero or nonzero? 11–73C Air is flowing past a nonsymmetrical airfoil at zero angle of attack. Will the (a) lift and (b) drag acting on the airfoil be zero or nonzero? 11–74C Air is flowing past a symmetrical airfoil at an angle of attack of 5°. Will the (a) lift and (b) drag acting on the airfoil be zero or nonzero? 11–75C What is stall? What causes an airfoil to stall? Why are commercial aircraft not allowed to fly at conditions near stall? 11–76C Both the lift and the drag of an airfoil increase with an increase in the angle of attack. In general, which increases at a higher rate, the lift or the drag? 11–77C Why are flaps used at the leading and trailing edges of the wings of large aircraft during takeoff and landing? Can an aircraft take off or land without them?
θ
11–78C 2m 40 km/h 0.2 m
FIGURE P11–69 11–70 One of the popular demonstrations in science museums involves the suspension of a ping-pong ball by an upward air jet. Children are amused by the ball always coming back to the center when it is pushed by a finger to the side of the jet. Explain this phenomenon using the Bernoulli equation. Also determine the velocity of air if the ball has a mass of 2.6 g and a diameter of 3.8 cm. Assume air is at 1 atm and 25°C.
How do flaps affect the lift and the drag of wings?
11–79C What is the effect of wing tip vortices (the air circulation from the lower part of the wings to the upper part) on the drag and the lift? 11–80C What is induced drag on wings? Can induced drag be minimized by using long and narrow wings or short and wide wings? 11–81C Air is flowing past a spherical ball. Is the lift exerted on the ball zero or nonzero? Answer the same question if the ball is spinning. 11–82 A tennis ball with a mass of 57 g and a diameter of 6.4 cm is hit with an initial velocity of 92 km/h and a backspin 4200 rpm
92 km/h
Air jet Ball
FIGURE P11–70
FIGURE P11–82
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of 4200 rpm. Determine if the ball will fall or rise under the combined effect of gravity and lift due to spinning shortly after hitting. Assume air is at 1 atm and 25°C.
CL # 0.45 30 m2
11–83 Consider an aircraft, which takes off at 190 km/h when it is fully loaded. If the weight of the aircraft is increased by 20 percent as a result of overloading, determine the speed at which the overloaded aircraft will take off.
2800 kg
Answer: 208 km/h
11–84 Consider an airplane whose takeoff speed is 220 km/h and that takes 15 s to take off at sea level. For an airport at an elevation of 1600 m (such as Denver), determine (a) the takeoff speed, (b) the takeoff time, and (c) the additional runway length required for this airplane. Assume constant acceleration for both cases.
FIGURE P11–88 of this airplane while cruising at an altitude of 4000 m at a constant speed of 280 km/h and generating 190 kW of power. 11–90 The NACA 64(l)–412 airfoil has a lift-to-drag ratio of 50 at 0° angle of attack, as shown in Fig. 11–43. At what angle of attack will this ratio increase to 80? 11–91 Consider a light plane that has a total weight of 15,000 N and a wing area of 46 m2 and whose wings resemble the NACA 23012 airfoil with no flaps. Using data from Fig. 11–45, determine the takeoff speed at an angle of attack of 5° at sea level. Also determine the stall speed. Answers: 94 km/h, 67.4 km/h
220 km/h
FIGURE P11–84 11–85E An airplane is consuming fuel at a rate of 5 gal/min when cruising at a constant altitude of 10,000 ft at constant speed. Assuming the drag coefficient and the engine efficiency to remain the same, determine the rate of fuel consumption at an altitude of 30,000 ft at the same speed. 11–86 A jumbo jet airplane has a mass of about 400,000 kg when fully loaded with over 400 passengers and takes off at a speed of 250 km/h. Determine the takeoff speed when the airplane has 100 empty seats. Assume each passenger with luggage is 140 kg and the wing and flap settings are maintained the same. Answer: 246 km/h 11–87
Reconsider Prob. 11–86. Using EES (or other) software, investigate the effect of passenger count on the takeoff speed of the aircraft. Let the number of passengers vary from 0 to 500 in increments of 50. Tabulate and plot the results.
11–92
An airplane has a mass of 50,000 kg, a wing area of 300 m2, a maximum lift coefficient of 3.2, and a cruising drag coefficient of 0.03 at an altitude of 12,000 m. Determine (a) the takeoff speed at sea level, assuming it is 20 percent over the stall speed, and (b) the thrust that the engines must deliver for a cruising speed of 700 km/h. 11–93E A 2.4-in-diameter smooth ball rotating at 500 rpm is dropped in a water stream at 60°F flowing at 4 ft/s. Determine the lift and the drag force acting on the ball when it is first dropped in the water.
Review Problems 11–94 An automotive engine can be approximated as a 0.4m-high, 0.60-m-wide, and 0.7-m-long rectangular block. The Air 85 km/h 15°C Engine block
11–88 A small aircraft has a wing area of 30 m2, a lift coefficient of 0.45 at takeoff settings, and a total mass of 2800 kg. Determine (a) the takeoff speed of this aircraft at sea level at standard atmospheric conditions, (b) the wing loading, and (c) the required power to maintain a constant cruising speed of 300 km/h for a cruising drag coefficient of 0.035. 11–89 A small airplane has a total mass of 1800 kg and a wing area of 42 m2. Determine the lift and drag coefficients
FIGURE P11–94
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ambient air is at 1 atm and 15°C. Determine the drag force acting on the bottom surface of the engine block as the car travels at a velocity of 85 km/h. Assume the flow to be turbulent over the entire surface because of the constant agitation of the engine block. Answer: 0.65 N 11–95
Calculate the thickness of the boundary layer during flow over a 2.5-m-long flat plate at intervals of 25 cm and plot the boundary layer over the plate for the flow of (a) air, (b) water, and (c) engine oil at 1 atm and 20°C at an upstream velocity of 3 m/s. 11–96E The passenger compartment of a minivan traveling at 60 mi/h in ambient air at 1 atm and 80°F can be modeled as a 3.2-ft-high, 6-ft-wide, and 11-ft-long rectangular box. The airflow over the exterior surfaces can be assumed to be turbulent because of the intense vibrations involved. Determine the drag force acting on the top and the two side surfaces of the van and the power required to overcome it.
crete block (density # 2300 kg/m3) by two 5-cm-diameter, 4-m-high (exposed part) poles, as shown in Fig. P11–98. If the sign is to withstand 150 km/h winds from any direction, determine (a) the maximum drag force on the panel, (b) the drag force acting on the poles, and (c) the minimum length L of the concrete block for the panel to resist the winds. Take the density of air to be 1.30 kg/m3. 11–99 A plastic boat whose bottom surface can be approximated as a 1.5-m-wide, 2-m-long flat surface is to move through water at 15°C at speeds up to 30 km/h. Determine the friction drag exerted on the boat by water and the power needed to overcome it.
30 km/h
Air 60 mi/h 80°F
FIGURE P11–99
11–100
FIGURE P11–96E 11–97 A 1-m-external-diameter spherical tank is located outdoors at 1 atm and 25°C and is subjected to winds at 35 km/h. Determine the drag force exerted on it by the wind. Answer: 3.5 N
11–98 A 2-m-high, 4-m-wide rectangular advertisement panel is attached to a 4-m-wide, 0.15-m-high rectangular con-
4m 2m
4m Concrete
4m 0.15 m
FIGURE P11–98
Reconsider Prob. 11–99. Using EES (or other) software, investigate the effect of boat speed on the drag force acting on the bottom surface of the boat, and the power needed to overcome it. Let the boat speed vary from 0 to 100 km/h in increments of 10 km/h. Tabulate and plot the results. 11–101E
A commercial airplane has a total mass of 150,000 lbm and a wing planform area of 1800 ft2. The plane has a cruising speed of 550 mi/h and a cruising altitude of 38,000 ft where the air density is 0.0208 lbm/ft3. The plane has double-slotted flaps for use during takeoff and landing, but it cruises with all flaps retracted. Assuming the lift and drag characteristics of the wings can be approximated by NACA 23012, determine (a) the minimum safe speed for takeoff and landing with and without extending the flaps, (b) the angle of attack to cruise steadily at the cruising altitude, and (c) the power that needs to be supplied to provide enough thrust to overcome drag. Take the air density on the ground to be 0.075 lbm/ft3. 11–102 An 8-cm-diameter smooth ball has a velocity of 36 km/h during a typical hit. Determine the percent increase in the drag coefficient if the ball is given a spin of 3500 rpm in air at 1 atm and 25°C. 11–103 A paratrooper and his 8-m-diameter parachute weigh 950 N. Taking the average air density to be 1.2 kg/m3, determine the terminal velocity of the paratrooper. Answer: 4.9 m/s
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609 CHAPTER 11
observing when the curve becomes linear. During such an experiment a 3-mm-diameter glass ball (r # 2500 kg/m3) is dropped into a fluid whose density is 875 kg/m3, and the terminal velocity is measured to be 0.12 m/s. Disregarding the wall effects, determine the viscosity of the fluid.
8m
11–106 During an experiment, three aluminum balls (rs # 2600 kg/m3) having diameters 2, 4, and 10 mm, respectively, are dropped into a tank filled with glycerin at 22°C (rf # 1274 kg/m3 and m # 1 kg/m · s). The terminal settling velocities of the balls are measured to be 3.2, 12.8, and 60.4 mm/s, respectively. Compare these values with the velocities predicted by Stokes law for drag force FD # 3pmDV, which is valid for very low Reynolds numbers (Re (( 1). Determine the error involved for each case and assess the accuracy of Stokes law.
950 N
11–107 Repeat Prob. 11–106 by considering the general form of Stokes law expressed as FD # 3pmDV % (9p/16)rsV 2D2.
FIGURE P11–103 11–104 A 17,000-kg tractor-trailer rig has a frontal area of 9.2 m2, a drag coefficient of 0.96, a rolling resistance coefficient of 0.05 (multiplying the weight of a vehicle by the rolling resistance coefficient gives the rolling resistance), a bearing friction resistance of 350 N, and a maximum speed of 110 km/h on a level road during steady cruising in calm weather with an air density of 1.25 kg/m3. Now a fairing is installed to the front of the rig to suppress separation and to streamline the flow to the top surface, and the drag coefficient is reduced to 0.76. Determine the maximum speed of the rig with the fairing. Answer: 133 km/h 11–105 Stokes law can be used to determine the viscosity of a fluid by dropping a spherical object in it and measuring the terminal velocity of the object in that fluid. This can be done by plotting the distance traveled against time and
Glass ball
FIGURE P11–105
11–108 A small aluminum ball with D # 2 mm and rs # 2700 kg/m3 is dropped into a large container filled with oil at 40°C (rf # 876 kg/m3 and m # 0.2177 kg/m · s). The Reynolds number is expected to be low and thus Stokes law for drag force FD # 3pmDV to be applicable. Show that the variation of velocity with time can be expressed as V # (a/b)(1 $ e$bt) where a # g(1 $ rf /rs) and b # 18m/(rs D2). Plot the variation of velocity with time, and calculate the time it takes for the ball to reach 99 percent of its terminal velocity. 11–109
Engine oil at 40°C is flowing over a long flat plate with a velocity of 4 m/s. Determine the distance xcr from the leading edge of the plate where the flow becomes turbulent, and calculate and plot the thickness of the boundary layer over a length of 2xcr.
Design and Essay Problems 11–110 Write a report on the history of the reduction of the drag coefficients of cars and obtain the drag coefficient data for some recent car models from the catalogs of car manufacturers. 11–111 Write a report on the flaps used at the leading and trailing edges of the wings of large commercial aircraft. Discuss how the flaps affect the drag and lift coefficients during takeoff and landing.
0.12 m/s
11–112 Large commercial airplanes cruise at high altitudes (up to about 40,000 ft) to save fuel. Discuss how flying at high altitudes reduces drag and saves fuel. Also discuss why small planes fly at relatively low altitudes. 11–113 Many drivers turn off their air conditioners and roll down the car windows in hopes of saving fuel. But it is claimed that this apparent “free cooling” actually increases the fuel consumption of the car. Investigate this matter and write a report on which practice will save gasoline under what conditions.
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CHAPTER
12
COMPRESSIBLE FLOW
F
or the most part, we have limited our consideration so far to flows for which density variations and thus compressibility effects are negligible. In this chapter we lift this limitation and consider flows that involve significant changes in density. Such flows are called compressible flows, and they are frequently encountered in devices that involve the flow of gases at very high speeds. Compressible flow combines fluid dynamics and thermodynamics in that both are absolutely necessary to the development of the required theoretical background. In this chapter, we develop the general relations associated with compressible flows for an ideal gas with constant specific heats. We start this chapter by introducing the concepts of stagnation state, speed of sound, and Mach number for compressible flows. The relationships between the static and stagnation fluid properties are developed for isentropic flows of ideal gases, and they are expressed as functions of specific heat ratios and the Mach number. The effects of area changes for onedimensional isentropic subsonic and supersonic flows are discussed. These effects are illustrated by considering the isentropic flow through converging and converging–diverging nozzles. The concept of shock waves and the variation of flow properties across normal and oblique shock waves are discussed. Finally, we consider the effects of friction and heat transfer on compressible flows and develop relations for property changes.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
■
Appreciate the consequences of compressibility in gas flow Understand why a nozzle must have a diverging section to accelerate a gas to supersonic speeds Predict the occurrence of shocks and calculate property changes across a shock wave Understand the effects of friction and heat transfer on compressible flows
611
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612 FLUID MECHANICS
12–1
■
STAGNATION PROPERTIES
When analyzing control volumes, we find it very convenient to combine the internal energy and the flow energy of a fluid into a single term, enthalpy, defined per unit mass as h ! u " P/r. Whenever the kinetic and potential energies of the fluid are negligible, as is often the case, the enthalpy represents the total energy of a fluid. For high-speed flows, such as those encountered in jet engines (Fig. 12–1), the potential energy of the fluid is still negligible, but the kinetic energy is not. In such cases, it is convenient to combine the enthalpy and the kinetic energy of the fluid into a single term called stagnation (or total) enthalpy h0, defined per unit mass as
(a)
h0 ! h "
(b)
FIGURE 12–1 Aircraft and jet engines involve high speeds, and thus the kinetic energy term should always be considered when analyzing them. (a) Photo courtesy NASA, http://lisar.larc.nasa. gov/IMAGES/SMALL/EL-1999-00108.jpeg, and (b) Figure courtesy Pratt and Whitney. Used by permission.
V2 2
(kJ/kg)
(12–1)
When the potential energy of the fluid is negligible, the stagnation enthalpy represents the total energy of a flowing fluid stream per unit mass. Thus it simplifies the thermodynamic analysis of high-speed flows. Throughout this chapter the ordinary enthalpy h is referred to as the static enthalpy, whenever necessary, to distinguish it from the stagnation enthalpy. Notice that the stagnation enthalpy is a combination property of a fluid, just like the static enthalpy, and these two enthalpies become identical when the kinetic energy of the fluid is negligible. Consider the steady flow of a fluid through a duct such as a nozzle, diffuser, or some other flow passage where the flow takes place adiabatically and with no shaft or electrical work, as shown in Fig. 12–2. Assuming the fluid experiences little or no change in . its elevation and its potential energy, . the energy balance relation (E in ! E out) for this single-stream steady-flow device reduces to h1 "
V 21 V 22 ! h2 " 2 2
(12–2)
h01 ! h02
(12–3)
or That is, in the absence of any heat and work interactions and any changes in potential energy, the stagnation enthalpy of a fluid remains constant during a steady-flow process. Flows through nozzles and diffusers usually satisfy these conditions, and any increase in fluid velocity in these devices creates an equivalent decrease in the static enthalpy of the fluid. If the fluid were brought to a complete stop, then the velocity at state 2 would be zero and Eq. 12–2 would become h1 V1 h01
Control volume
h2 V2 h02 = h 01
FIGURE 12–2 Steady flow of a fluid through an adiabatic duct.
h1 "
V 21 ! h2 ! h02 2
Thus the stagnation enthalpy represents the enthalpy of a fluid when it is brought to rest adiabatically. During a stagnation process, the kinetic energy of a fluid is converted to enthalpy (internal energy " flow energy), which results in an increase in the fluid temperature and pressure (Fig. 12–3). The properties of a fluid at the stagnation state are called stagnation properties (stagnation temperature,
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613 CHAPTER 12
stagnation pressure, stagnation density, etc.). The stagnation state and the stagnation properties are indicated by the subscript 0. The stagnation state is called the isentropic stagnation state when the stagnation process is reversible as well as adiabatic (i.e., isentropic). The entropy of a fluid remains constant during an isentropic stagnation process. The actual (irreversible) and isentropic stagnation processes are shown on an h-s diagram in Fig. 12–4. Notice that the stagnation enthalpy of the fluid (and the stagnation temperature if the fluid is an ideal gas) is the same for both cases. However, the actual stagnation pressure is lower than the isentropic stagnation pressure since entropy increases during the actual stagnation process as a result of fluid friction. The stagnation processes are often approximated to be isentropic, and isentropic stagnation properties are simply referred to as stagnation properties. When the fluid is approximated as an ideal gas with constant specific heats, its enthalpy can be replaced by cpT and Eq. 12–1 can be expressed as V2 2
© Reprinted with special permission of King Features Syndicate.
T0 ! T "
V2 2cp
(12–4)
Here T0 is called the stagnation (or total) temperature, and it represents the temperature an ideal gas attains when it is brought to rest adiabatically. The term V 2/2cp corresponds to the temperature rise during such a process and is called the dynamic temperature. For example, the dynamic temperature of air flowing at 100 m/s is (100 m/s)2/(2 # 1.005 kJ/kg · K) ! 5.0 K. Therefore, when air at 300 K and 100 m/s is brought to rest adiabatically (at the tip of a temperature probe, for example), its temperature rises to the stagnation value of 305 K (Fig. 12–5). Note that for low-speed flows, the stagnation and static (or ordinary) temperatures are practically the same. But for high-speed flows, the temperature measured by a stationary probe placed in the fluid (the stagnation temperature) may be significantly higher than the static temperature of the fluid. The pressure a fluid attains when brought to rest isentropically is called the stagnation pressure P0. For ideal gases with constant specific heats, P0 is related to the static pressure of the fluid by T0 k/(k$1) P0 !a b P T
(12–5)
By noting that r ! 1/v and using the isentropic relation Pv k ! P0v k0, the ratio of the stagnation density to static density can be expressed as r0 T0 1/(k$1) !a b r T
(12–6)
When stagnation enthalpies are used, there .is no need . to refer explicitly to kinetic energy. Then the energy balance Ein ! Eout for a single-stream, steady-flow device can be expressed as qin " win " (h01 " gz 1) ! qout " wout " (h02 " gz 2)
(12–7)
0,
ac
t
P
Isentropic stagnation state
or
0
h P
cpT0 ! cpT "
FIGURE 12–3 Kinetic energy is converted to enthalpy during a stagnation process.
h0
V2 2
Actual stagnation state
P
h Actual state s
FIGURE 12–4 The actual state, actual stagnation state, and isentropic stagnation state of a fluid on an h-s diagram.
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614 FLUID MECHANICS
where h01 and h02 are the stagnation enthalpies at states 1 and 2, respectively. When the fluid is an ideal gas with constant specific heats, Eq. 12–7 becomes
Temperature rise during stagnation
(qin $ qout) " (win $ wout) ! cp(T02 $ T01) " g(z 2 $ z 1) 305 K 300 K
(12–8)
where T01 and T02 are the stagnation temperatures. Notice that kinetic energy terms do not explicitly appear in Eqs. 12–7 and 12–8, but the stagnation enthalpy terms account for their contribution.
AIR 100 m/s
EXAMPLE 12–1
FIGURE 12–5 The temperature of an ideal gas flowing at a velocity V rises by V 2/2cp when it is brought to a complete stop.
Diffuser
Compressor
T1 = 255.7 K
Aircraft engine
P1 = 54.05 kPa V 1 = 250 m/s
1
01
02
FIGURE 12–6 Schematic for Example 12–1.
Compression of High-Speed Air in an Aircraft
An aircraft is flying at a cruising speed of 250 m/s at an altitude of 5000 m where the atmospheric pressure is 54.05 kPa and the ambient air temperature is 255.7 K. The ambient air is first decelerated in a diffuser before it enters the compressor (Fig. 12–6). Assuming both the diffuser and the compressor to be isentropic, determine (a) the stagnation pressure at the compressor inlet and (b) the required compressor work per unit mass if the stagnation pressure ratio of the compressor is 8.
SOLUTION High-speed air enters the diffuser and the compressor of an aircraft. The stagnation pressure of the air and the compressor work input are to be determined. Assumptions 1 Both the diffuser and the compressor are isentropic. 2 Air is an ideal gas with constant specific heats at room temperature. Properties The constant-pressure specific heat cp and the specific heat ratio k of air at room temperature are cp ! 1.005 kJ/kg % K
and
k ! 1.4
Analysis (a) Under isentropic conditions, the stagnation pressure at the compressor inlet (diffuser exit) can be determined from Eq. 12–5. However, first we need to find the stagnation temperature T01 at the compressor inlet. Under the stated assumptions, T01 can be determined from Eq. 12–4 to be
T01 ! T1 "
1 kJ/kg V 21 (250 m/s)2 ! 255.7 K " a b 2cp (2)(1.005 kJ/kg % K) 1000 m2/s2
! 286.8 K Then from Eq. 12–5,
T01 k/(k$1) 286.8 K 1.4/(1.4$1) b ! (54.05 kPa)a b T1 255.7 K
P01 ! P1 a
! 80.77 kPa
That is, the temperature of air would increase by 31.1°C and the pressure by 26.72 kPa as air is decelerated from 250 m/s to zero velocity. These increases in the temperature and pressure of air are due to the conversion of the kinetic energy into enthalpy. (b) To determine the compressor work, we need to know the stagnation temperature of air at the compressor exit T02. The stagnation pressure ratio across the compressor P02/P01 is specified to be 8. Since the compression process is assumed to be isentropic, T02 can be determined from the idealgas isentropic relation (Eq. 12–5):
P02 (k$1)/k b ! (286.8 K)(8)(1.4$1)/1.4 ! 519.5 K P01
T02 ! T01 a
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615 CHAPTER 12
Disregarding potential energy changes and heat transfer, the compressor work per unit mass of air is determined from Eq. 12–8:
win ! cp(T02 $ T01) ! (1.005 kJ/kg % K)(519.5 K $ 286.8 K) ! 233.9 kJ/kg Thus the work supplied to the compressor is 233.9 kJ/kg. Discussion Notice that using stagnation properties automatically accounts for any changes in the kinetic energy of a fluid stream.
12–2
■
SPEED OF SOUND AND MACH NUMBER
An important parameter in the study of compressible flow is the speed of sound (or the sonic speed), which is the speed at which an infinitesimally small pressure wave travels through a medium. The pressure wave may be caused by a small disturbance, which creates a slight rise in local pressure. To obtain a relation for the speed of sound in a medium, consider a duct that is filled with a fluid at rest, as shown in Fig. 12–7. A piston fitted in the duct is now moved to the right with a constant incremental velocity dV, creating a sonic wave. The wave front moves to the right through the fluid at the speed of sound c and separates the moving fluid adjacent to the piston from the fluid still at rest. The fluid to the left of the wave front experiences an incremental change in its thermodynamic properties, while the fluid on the right of the wave front maintains its original thermodynamic properties, as shown in Fig. 12–7. To simplify the analysis, consider a control volume that encloses the wave front and moves with it, as shown in Fig. 12–8. To an observer traveling with the wave front, the fluid to the right will appear to be moving toward the wave front with a speed of c and the fluid to the left to be moving away from the wave front with a speed of c $ dV. Of course, the observer will think the control volume that encloses the wave front (and herself or himself) is stationary, and the observer will be witnessing a steady-flow process. The mass balance for this single-stream, steady-flow process can be expressed as # # mright ! mleft
dV
(a)
No heat or work crosses the boundaries of the control volume during this steady-flow process, and the potential energy change can be neglected. Then the steady-flow energy balance ein ! eout becomes (c $ dV)2 c2 h " ! h " dh " 2 2
c
h Stationary P fluid r
dV 0
x
P
P + dP P x
FIGURE 12–7 Propagation of a small pressure wave along a duct.
Control volume traveling with the wave front
rAc ! (r " dr)A(c $ dV)
c dr $ r dV ! 0
h + dh P + dP r + dr
V
or By canceling the cross-sectional (or flow) area A and neglecting the higherorder terms, this equation reduces to
Moving wave front
Piston
h + dh P + dP r + dr
c – dV
c
h P r
FIGURE 12–8 Control volume moving with the small pressure wave along a duct.
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616 FLUID MECHANICS
which yields dh $ c dV ! 0
(b)
dV 2.
where we have neglected the second-order term The amplitude of the ordinary sonic wave is very small and does not cause any appreciable change in the pressure and temperature of the fluid. Therefore, the propagation of a sonic wave is not only adiabatic but also very nearly isentropic. Then the thermodynamic relation T ds ! dh $ dP/r (see Çengel, 2002) reduces to 0 dP T¡ ds ! dh $ r
or dP r
dh !
(c)
Combining Eqs. a, b, and c yields the desired expression for the speed of sound as c2 !
dP dr
at s ! constant
or c2 ! a
&P b &r s
(12–9)
It is left as an exercise for the reader to show, by using thermodynamic property relations (see Ref. 1) that Eq. 12–9 can also be written as &P b &r T
c 2 ! ka
(12–10)
where k is the specific heat ratio of the fluid. Note that the speed of sound in a fluid is a function of the thermodynamic properties of that fluid. When the fluid is an ideal gas (P ! rRT), the differentiation in Eq. 12–10 can easily be performed to yield &(rRT) &P b ! kc d ! kRT &r T &r T
c 2 ! ka AIR
HELIUM
284 m/s
200 K
347 m/s
300 K
832 m/s
1019 m/s
1000 K 634 m/s
1861 m/s
or
c ! 2kRT
Noting that the gas constant R has a fixed value for a specified ideal gas and the specific heat ratio k of an ideal gas is, at most, a function of temperature, we see that the speed of sound in a specified ideal gas is a function of temperature alone (Fig. 12–9). A second important parameter in the analysis of compressible fluid flow is the Mach number Ma, named after the Austrian physicist Ernst Mach (1838–1916). It is the ratio of the actual velocity of the fluid (or an object in still fluid) to the speed of sound in the same fluid at the same state: Ma !
FIGURE 12–9 The speed of sound changes with temperature and varies with the fluid.
(12–11)
V c
(12–12)
Note that the Mach number depends on the speed of sound, which depends on the state of the fluid. Therefore, the Mach number of an aircraft cruising
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617 CHAPTER 12
at constant velocity in still air may be different at different locations (Fig. 12–10). Fluid flow regimes are often described in terms of the flow Mach number. The flow is called sonic when Ma ! 1, subsonic when Ma ' 1, supersonic when Ma ( 1, hypersonic when Ma (( 1, and transonic when Ma ! 1.
EXAMPLE 12–2
Mach Number of Air Entering a Diffuser
Air enters a diffuser shown in Fig. 12–11 with a velocity of 200 m/s. Determine (a) the speed of sound and (b) the Mach number at the diffuser inlet when the air temperature is 30°C.
SOLUTION Air enters a diffuser with a high velocity. The speed of sound and the Mach number are to be determined at the diffuser inlet. Assumption Air at specified conditions behaves as an ideal gas. Properties The gas constant of air is R ! 0.287 kJ/kg · K, and its specific heat ratio at 30°C is 1.4. Analysis We note that the speed of sound in a gas varies with temperature, which is given to be 30°C.
AIR 200 K
V = 320 m/s Ma = 1.13
V = 320 m/s
AIR 300 K
Ma = 0.92
FIGURE 12–10 The Mach number can be different at different temperatures even if the flight speed is the same.
AIR V = 200 m/s T = 30°C
Diffuser
(a) The speed of sound in air at 30°C is determined from Eq. 12–11 to be
c ! 2kRT !
B
(1.4)(0.287 kJ/kg % K)(303 K)a
(b) Then the Mach number becomes
Ma !
1000 m2/s2 b ! 349 m/s 1 kJ/kg
FIGURE 12–11 Schematic for Example 12–2.
V 200 m/s ! ! 0.573 c 349 m/s
Discussion The flow at the diffuser inlet is subsonic since Ma ' 1.
12–3
■
ONE-DIMENSIONAL ISENTROPIC FLOW
During fluid flow through many devices such as nozzles, diffusers, and turbine blade passages, flow quantities vary primarily in the flow direction only, and the flow can be approximated as one-dimensional isentropic flow with good accuracy. Therefore, it merits special consideration. Before presenting a formal discussion of one-dimensional isentropic flow, we illustrate some important aspects of it with an example.
EXAMPLE 12–3
Gas Flow through a Converging–Diverging Duct
Carbon dioxide flows steadily through a varying cross-sectional area duct such as a nozzle shown in Fig. 12–12 at a mass flow rate of 3 kg/s. The carbon dioxide enters the duct at a pressure of 1400 kPa and 200°C with a low velocity, and it expands in the nozzle to a pressure of 200 kPa. The duct is designed so that the flow can be approximated as isentropic. Determine the density, velocity, flow area, and Mach number at each location along the duct that corresponds to a pressure drop of 200 kPa.
Stagnation region: m⋅ ! 1400 kPa 3.00 kg/s 200°C CO2
1400
1000 767 P, kPa
200
FIGURE 12–12 Schematic for Example 12–3.
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618 FLUID MECHANICS
SOLUTION Carbon dioxide enters a varying cross-sectional area duct at specified conditions. The flow properties are to be determined along the duct. Assumptions 1 Carbon dioxide is an ideal gas with constant specific heats at room temperature. 2 Flow through the duct is steady, one-dimensional, and isentropic. Properties For simplicity we use cp ! 0.846 kJ/kg · K and k ! 1.289 throughout the calculations, which are the constant-pressure specific heat and specific heat ratio values of carbon dioxide at room temperatures. The gas constant of carbon dioxide is R ! 0.1889 kJ/kg · K. Analysis We note that the inlet temperature is nearly equal to the stagnation temperature since the inlet velocity is small. The flow is isentropic, and thus the stagnation temperature and pressure throughout the duct remain constant. Therefore, T0 " T1 ! 200)C ! 473 K and
P0 " P1 ! 1400 kPa To illustrate the solution procedure, we calculate the desired properties at the location where the pressure is 1200 kPa, the first location that corresponds to a pressure drop of 200 kPa. From Eq. 12–5,
P (k$1)/k 1200 kPa (1.289$1)/1.289 T ! T0 a b ! (473 K) a ! 457 K b P0 1400 kPa
From Eq. 12–4,
V ! 22cp(T0 $ T) !
B
1000 m2/s3 b 1 kJ/kg
2(0.846 kJ/kg % K)(473 K $ 457 K)a
! 164.5 m/s From the ideal-gas relation,
r!
1200 kPa P ! ! 13.9 kg/m3 RT (0.1889 kPa % m3/kg % K)(457 K)
From the mass flow rate relation,
# 3 kg/s m ! ! 13.1 # 10 $4 m2 ! 13.1 cm2 A! rV (13.9 kg/m3)(164.5 m/s)
From Eqs. 12–11 and 12–12,
c ! 2kRT ! Ma !
B
(1.289)(0.1889 kJ/kg % K)(457 K)a
V 164.5 m/s ! ! 0.493 c 333.6 m/s
1000 m2/s2 b ! 333.6 m/s 1 kJ/kg
The results for the other pressure steps are summarized in Table 12–1 and are plotted in Fig. 12–13. Discussion Note that as the pressure decreases, the temperature and speed of sound decrease while the fluid velocity and Mach number increase in the flow direction. The density decreases slowly at first and rapidly later as the fluid velocity increases.
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619 CHAPTER 12
TA B L E 1 2 – 1 Variation of fluid properties in flow direction in the duct described in Example 12–3 for m% ! 3 kg/s ! constant P, kPa
T, K
V, m/s
r, kg/m3
c, m/s
A, cm2
Ma
1400 1200 1000 800 767* 600 400 200
473 457 439 417 413 391 357 306
0 164.5 240.7 306.6 317.2 371.4 441.9 530.9
15.7 13.9 12.1 10.1 9.82 8.12 5.93 3.46
339.4 333.6 326.9 318.8 317.2 308.7 295.0 272.9
* 13.1 10.3 9.64 9.63 10.0 11.5 16.3
0 0.493 0.736 0.962 1.000 1.203 1.498 1.946
* 767 kPa is the critical pressure where the local Mach number is unity.
Flow direction A, Ma, r, T, V
A T r
Ma V 1400
1200
1000
800 P, kPa
600
400
200
We note from Example 12–3 that the flow area decreases with decreasing pressure down to a critical-pressure value where the Mach number is unity, and then it begins to increase with further reductions in pressure. The Mach number is unity at the location of smallest flow area, called the throat (Fig. 12–14). Note that the velocity of the fluid keeps increasing after passing the throat although the flow area increases rapidly in that region. This increase in velocity past the throat is due to the rapid decrease in the fluid density. The flow area of the duct considered in this example first decreases and then increases. Such ducts are called converging–diverging nozzles. These nozzles are used to accelerate gases to supersonic speeds and should not be confused with Venturi nozzles, which are used strictly for incompressible flow. The first use of such a nozzle occurred in 1893 in a steam turbine designed by a Swedish engineer, Carl G. B. de Laval (1845–1913), and therefore converging–diverging nozzles are often called Laval nozzles.
FIGURE 12–13 Variation of normalized fluid properties and cross-sectional area along a duct as the pressure drops from 1400 to 200 kPa.
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620 FLUID MECHANICS Throat Fluid
Converging nozzle Throat Fluid
FIGURE 12–14 The cross section of a nozzle at the smallest flow area is called the throat.
Converging–diverging nozzle
Variation of Fluid Velocity with Flow Area
It is clear from Example 12–3 that the couplings among the velocity, density, and flow areas for isentropic duct flow are rather complex. In the remainder of this section we investigate these couplings more thoroughly, and we develop relations for the variation of static-to-stagnation property ratios with the Mach number for pressure, temperature, and density. We begin our investigation by seeking relationships among the pressure, temperature, density, velocity, flow area, and Mach number for one-dimensional isentropic flow. Consider the mass balance for a steady-flow process: # m ! rAV ! constant
Differentiating and dividing the resultant equation by the mass flow rate, we obtain dr dA dV " " !0 r A V CONSERVATION OF ENERGY (steady flow, w = 0, q = 0, ∆pe = 0) h1 + or h+
V 12 V2 = h2 + 2 2 2
V2
2 Differentiate,
= constant
dh + V dV = 0 Also,
0 (isentropic) T ds = dh – v dP 1 dh = v dP = r dP Substitute, dP r + V dV = 0
FIGURE 12–15 Derivation of the differential form of the energy equation for steady isentropic flow.
(12–13)
Neglecting the potential energy, the energy balance for an isentropic flow with no work interactions can be expressed in the differential form as (Fig. 12–15) dP " V dV ! 0 r
(12–14)
This relation is also the differential form of Bernoulli’s equation when changes in potential energy are negligible, which is a form of the conservation of momentum principle for steady-flow control volumes. Combining Eqs. 12–13 and 12–14 gives dr dA dP 1 a 2$ b ! r A dP V
(12–15)
Rearranging Eq. 12–9 as (&r/&P)s ! 1/c2 and substituting into Eq. 12–15 yield dA dP ! (1 $ Ma2) A rV 2
(12–16)
This is an important relation for isentropic flow in ducts since it describes the variation of pressure with flow area. We note that A, r, and V are positive quantities. For subsonic flow (Ma ' 1), the term 1 $ Ma2 is positive;
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621 CHAPTER 12
and thus dA and dP must have the same sign. That is, the pressure of the fluid must increase as the flow area of the duct increases and must decrease as the flow area of the duct decreases. Thus, at subsonic velocities, the pressure decreases in converging ducts (subsonic nozzles) and increases in diverging ducts (subsonic diffusers). In supersonic flow (Ma ( 1), the term 1 $ Ma2 is negative, and thus dA and dP must have opposite signs. That is, the pressure of the fluid must increase as the flow area of the duct decreases and must decrease as the flow area of the duct increases. Thus, at supersonic velocities, the pressure decreases in diverging ducts (supersonic nozzles) and increases in converging ducts (supersonic diffusers). Another important relation for the isentropic flow of a fluid is obtained by substituting rV ! $dP/dV from Eq. 12–14 into Eq. 12–16: dA dV ! $ (1 $ Ma2) A V
(12–17)
This equation governs the shape of a nozzle or a diffuser in subsonic or supersonic isentropic flow. Noting that A and V are positive quantities, we conclude the following: For subsonic flow (Ma ' 1), For supersonic flow (Ma ( 1), For sonic flow (Ma ! 1),
dA '0 dV dA (0 dV dA !0 dV
Thus the proper shape of a nozzle depends on the highest velocity desired relative to the sonic velocity. To accelerate a fluid, we must use a converging nozzle at subsonic velocities and a diverging nozzle at supersonic velocities. The velocities encountered in most familiar applications are well below the sonic velocity, and thus it is natural that we visualize a nozzle as a converging duct. However, the highest velocity we can achieve by a converging nozzle is the sonic velocity, which occurs at the exit of the nozzle. If we extend the converging nozzle by further decreasing the flow area, in hopes of accelerating the fluid to supersonic velocities, as shown in Fig. 12–16, we are up for disappointment. Now the sonic velocity will occur at the exit of the converging extension, instead of the exit of the original nozzle, and the mass flow rate through the nozzle will decrease because of the reduced exit area. Based on Eq. 12–16, which is an expression of the conservation of mass and energy principles, we must add a diverging section to a converging nozzle to accelerate a fluid to supersonic velocities. The result is a converging– diverging nozzle. The fluid first passes through a subsonic (converging) section, where the Mach number increases as the flow area of the nozzle decreases, and then reaches the value of unity at the nozzle throat. The fluid continues to accelerate as it passes through a supersonic (diverging) section. . Noting that m ! rAV for steady flow, we see that the large decrease in density makes acceleration in the diverging section possible. An example of this type of flow is the flow of hot combustion gases through a nozzle in a gas turbine. The opposite process occurs in the engine inlet of a supersonic aircraft. The fluid is decelerated by passing it first through a supersonic diffuser,
A P 0, T0
MaA = 1 (sonic)
Converging nozzle
A P 0, T0
Converging nozzle
MaA < 1 B Ma B = 1 (sonic) Attachment
FIGURE 12–16 We cannot attain supersonic velocities by attaching a converging section to a converging nozzle. Doing so will only move the sonic cross section farther downstream and decrease the mass flow rate.
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622 FLUID MECHANICS
Ma
1
Supersonic nozzle
Supersonic diffuser (b) Supersonic flow
which has a flow area that decreases in the flow direction. Ideally, the flow reaches a Mach number of unity at the diffuser throat. The fluid is further decelerated in a subsonic diffuser, which has a flow area that increases in the flow direction, as shown in Fig. 12–17.
Property Relations for Isentropic Flow of Ideal Gases
Next we develop relations between the static properties and stagnation properties of an ideal gas in terms of the specific heat ratio k and the Mach number Ma. We assume the flow is isentropic and the gas has constant specific heats. The temperature T of an ideal gas anywhere in the flow is related to the stagnation temperature T0 through Eq. 12–4: T0 ! T "
V2 2cp
or T0 V2 !1" T 2cpT
Noting that cp ! kR/(k $ 1), c2 ! kRT, and Ma ! V/c, we see that V2 k$1 V2 k $ 1 V2 ! !a b 2!a b Ma2 2cpT 2[kR/(k $ 1)]T 2 2 c
Substituting yields
T0 k$1 !1" a b Ma2 T 2
which is the desired relation between T0 and T.
(12–18)
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623 CHAPTER 12
The ratio of the stagnation to static pressure is obtained by substituting Eq. 12–18 into Eq. 12–5: k/(k$1) P0 k$1 ! c1 " a b Ma2d P 2
(12–19)
The ratio of the stagnation to static density is obtained by substituting Eq. 12–18 into Eq. 12–6: 1/(k$1) r0 k$1 ! c1 " a b Ma2d r 2
(12–20)
Numerical values of T/T0, P/P0, and r/r0 are listed versus the Mach number in Table A–13 for k ! 1.4, which are very useful for practical compressible flow calculations involving air. The properties of a fluid at a location where the Mach number is unity (the throat) are called critical properties, and the ratios in Eqs. (12–18) through (12–20) are called critical ratios (Fig. 12–18). It is standard practice in the analysis of compressible flow to let the superscript asterisk (*) represent the critical values. Setting Ma ! 1 in Eqs. 12–18 through 12–20 yields T* 2 ! T0 k " 1
(12–21)
P* 2 k/(k$1) !a b P0 k"1
(12–22)
r* 2 1/(k$1) !a b r0 k"1
(12–23)
These ratios are evaluated for various values of k and are listed in Table 12–2. The critical properties of compressible flow should not be confused with the thermodynamic properties of substances at the critical point (such as the critical temperature Tc and critical pressure Pc).
Throat T0 P0 r0
Subsonic nozzle
T* P* r* (if Ma t = 1)
Throat T0 P0
Supersonic nozzle
r0 T * , P * , r* (Mat = 1)
FIGURE 12–18 When Mat ! 1, the properties at the nozzle throat become the critical properties.
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624 FLUID MECHANICS
TA B L E 1 2 – 2 The critical-pressure, critical-temperature, and critical-density ratios for isentropic flow of some ideal gases
P* P0 T* T0 r* r0
Superheated steam, k ! 1.3
Hot products of combustion, k ! 1.33
Air, k ! 1.4
Monatomic gases, k ! 1.667
0.5457
0.5404
0.5283
0.4871
0.8696
0.8584
0.8333
0.7499
0.6276
0.6295
0.6340
0.6495
EXAMPLE 12–4
Critical Temperature and Pressure in Gas Flow
Calculate the critical pressure and temperature of carbon dioxide for the flow conditions described in Example 12–3 (Fig. 12–19). P 0 = 1.4 MPa
CO2
T0 = 473 K P* T*
FIGURE 12–19 Schematic for Example 12–4.
SOLUTION For the flow discussed in Example 12–3, the critical pressure and temperature are to be calculated. Assumptions 1 The flow is steady, adiabatic, and one-dimensional. 2 Carbon dioxide is an ideal gas with constant specific heats. Properties The specific heat ratio of carbon dioxide at room temperature is k ! 1.289. Analysis The ratios of critical to stagnation temperature and pressure are determined to be T* 2 2 ! ! ! 0.87337 T0 k " 1 1.289 " 1 1.289/(1.289$1) P* 2 k/(k$1) 2 b b !a !a ! 0.5477 P0 k"1 1.289 " 1
Noting that the stagnation temperature and pressure are, from Example 12–3, T0 ! 473 K and P0 ! 1400 kPa, we see that the critical temperature and pressure in this case are
T* ! 0.87337T0 ! (0.87337)(473 K) ! 413 K P* ! 0.5477P0 ! (0.5477)(1400 kPa) ! 767 kPa Discussion Note that these values agree with those listed in Table 12–1, as expected. Also, property values other than these at the throat would indicate that the flow is not critical, and the Mach number is not unity.
12–4
■
ISENTROPIC FLOW THROUGH NOZZLES
Converging or converging–diverging nozzles are found in many engineering applications including steam and gas turbines, aircraft and spacecraft propulsion systems, and even industrial blasting nozzles and torch nozzles. In this section we consider the effects of back pressure (i.e., the pressure applied
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625 CHAPTER 12
at the nozzle discharge region) on the exit velocity, the mass flow rate, and the pressure distribution along the nozzle.
Converging Nozzles
Consider the subsonic flow through a converging nozzle as shown in Fig. 12–20. The nozzle inlet is attached to a reservoir at pressure Pr and temperature Tr. The reservoir is sufficiently large so that the nozzle inlet velocity is negligible. Since the fluid velocity in the reservoir is zero and the flow through the nozzle is approximated as isentropic, the stagnation pressure and stagnation temperature of the fluid at any cross section through the nozzle are equal to the reservoir pressure and temperature, respectively. Now we begin to reduce the back pressure and observe the resulting effects on the pressure distribution along the length of the nozzle, as shown in Fig. 12–20. If the back pressure Pb is equal to P1, which is equal to Pr, there is no flow and the pressure distribution is uniform along the nozzle. When the back pressure is reduced to P2, the exit plane pressure Pe also drops to P2. This causes the pressure along the nozzle to decrease in the flow direction. When the back pressure is reduced to P3 (! P*, which is the pressure required to increase the fluid velocity to the speed of sound at the exit plane or throat), the mass flow reaches a maximum value and the flow is said to be choked. Further reduction of the back pressure to level P4 or below does not result in additional changes in the pressure distribution, or anything else along the nozzle length. Under steady-flow conditions, the mass flow rate through the nozzle is constant and can be expressed as P k # m ! rAV ! a bA(Ma2kRT) ! PAMa RT B RT
Solving for T from Eq. 12–18 and for P from Eq. 12–19 and substituting, # m!
AMaP0 2k/(RT0) [1 " (k $ 1)Ma2/2](k"1)/[2(k$1)]
(12–24)
Thus the mass flow rate of a particular fluid through a nozzle is a function of the stagnation properties of the fluid, the flow area, and the Mach num. ber. Equation 12–24 is valid at any cross section, and thus m can be evaluated at any location along the length of the nozzle. For a specified flow area A and stagnation properties T0 and P0, the maximum mass flow rate can be determined by differentiating Eq. 12–24 with respect to Ma and setting the result equal to zero. It yields Ma ! 1. Since the only location in a nozzle where the Mach number can be unity is the location of minimum flow area (the throat), the mass flow rate through a nozzle is a maximum when Ma ! 1 at the throat. Denoting this area by A*, we obtain an expression for the maximum mass flow rate by substituting Ma ! 1 in Eq. 12–24: k 2 (k"1)/[2(k$1)] a b (12–25) B RT0 k " 1 Thus, for a particular ideal gas, the maximum mass flow rate through a nozzle with a given throat area is fixed by the stagnation pressure and temperature of the inlet flow. The flow rate can be controlled by changing the stagnation # m max ! A*P0
Reservoir Pr = P 0 Tr = T0
Pe Pb (Back pressure)
Vr = 0
x P/P0
1
1
2 P* P0
0
3 Lowest exit pressure
4 5
Pb = P0 Pb > P* Pb = P* Pb < P* Pb = 0 x
FIGURE 12–20 The effect of back pressure on the pressure distribution along a converging nozzle.
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626 FLUID MECHANICS m 5 m⋅ max
4
3
2
1 P* P0
Pe /P0
1.0
Pb P0
2
0
P Pe ! e b P*
1
1.0
P* 5 P0
pressure or temperature, and thus a converging nozzle can be used as a flowmeter. The flow rate can also be controlled, of course, by varying the throat area. This principle is very important for chemical processes, medical devices, flowmeters, and anywhere the mass flux of a gas must be known and controlled. . A plot of m versus Pb /P0 for a converging nozzle is shown in Fig. 12–21. Notice that the mass flow rate increases with decreasing Pb /P0, reaches a maximum at Pb ! P*, and remains constant for Pb /P0 values less than this critical ratio. Also illustrated on this figure is the effect of back pressure on the nozzle exit pressure Pe. We observe that
4
3
P* P0
1.0
Pb P0
FIGURE 12–21 The effect of back pressure Pb on the . mass flow rate m and the exit pressure Pe of a converging nozzle.
for Pb + P* for Pb ' P*
To summarize, for all back pressures lower than the critical pressure P*, the pressure at the exit plane of the converging nozzle Pe is equal to P*, the Mach number at the exit plane is unity, and the mass flow rate is the maximum (or choked) flow rate. Because the velocity of the flow is sonic at the throat for the maximum flow rate, a back pressure lower than the critical pressure cannot be sensed in the nozzle upstream flow and does not affect the flow rate. The effects of the stagnation temperature T0 and stagnation pressure P0 on the mass flow rate through a converging nozzle are illustrated in Fig. 12–22 where the mass flow rate is plotted against the static-to-stagnation pressure ratio at the throat Pt /P0. An increase in P0 (or a decrease of T0) will increase the mass flow rate through the converging nozzle; a decrease in P0 (or an increase in T0) will decrease it. We could also conclude this by carefully observing Eqs. 12–24 and 12–25. A relation for the variation of flow area A through the nozzle relative to throat area A* can be obtained by combining Eqs. 12–24 and 12–25 for the same mass flow rate and stagnation properties of a particular fluid. This yields (k"1)/[2(k$1)] A 1 2 k$1 ! ca b a1 " Ma2b d A* Ma k " 1 2
(12–26)
Table A–13 gives values of A/A* as a function of the Mach number for air (k ! 1.4). There is one value of A/A* for each value of the Mach number,
Mat = 1
Mat
0 0
EXAMPLE 12–8
The Point of Maximum Entropy on the Fanno Line
Show that the point of maximum entropy on the Fanno line (point a of Fig. 12–31) for the adiabatic steady flow of a fluid in a duct corresponds to the sonic velocity, Ma ! 1.
SOLUTION It is to be shown that the point of maximum entropy on the Fanno line for steady adiabatic flow corresponds to sonic velocity. Assumption The flow is steady, adiabatic, and one-dimensional. Analysis In the absence of any heat and work interactions and potential energy changes, the steady-flow energy equation reduces to h"
V2 ! constant 2
Differentiating yields
dh " V dV ! 0 For a very thin shock with negligible change of duct area across the shock, the steady-flow continuity (conservation of mass) equation can be expressed as
rV ! constant
s2 – s1 < 0
IMPOSSIBLE Subsonic flow before shock
Ma1 = 1 Supersonic flow Ma1 before shock
FIGURE 12–34 Entropy change across the normal shock.
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638 FLUID MECHANICS
Differentiating, we have
r dV " V dr ! 0 Solving for dV gives
dV ! $V
dr r
Combining this with the energy equation, we have
dr !0 r
dh $ V 2
which is the equation for the Fanno line in differential form. At point a (the point of maximum entropy) ds ! 0. Then from the second T ds relation (T ds ! dh $ v dP) we have dh ! v dP ! dP/r. Substituting yields
dr dP $ V2 !0 r r
at s ! constant
Solving for V, we have
&P 1/2 b &r s
V! a
which is the relation for the speed of sound, Eq. 12–9. Thus the proof is complete.
EXAMPLE 12–9
Shock Wave in a Converging–Diverging Nozzle
If the air flowing through the converging–diverging nozzle of Example 12–7 experiences a normal shock wave at the nozzle exit plane (Fig. 12–35), determine the following after the shock: (a) the stagnation pressure, static pressure, static temperature, and static density; (b) the entropy change across the shock; (c) the exit velocity; and (d) the mass flow rate through the nozzle. Assume steady, one-dimensional, and isentropic flow with k ! 1.4 from the nozzle inlet to the shock location.
Shock wave
m· = 2.858 kg/s
1
Ma1 = 2 P = 1.0 MPa 2 01 P1 = 0.1278 MPa T1 = 444.5 K r1 = 1.002 kg/m3
FIGURE 12–35 Schematic for Example 12–9.
SOLUTION Air flowing through a converging–diverging nozzle experiences a normal shock at the exit. The effect of the shock wave on various properties is to be determined. Assumptions 1 Air is an ideal gas with constant specific heats at room temperature. 2 Flow through the nozzle is steady, one-dimensional, and isentropic before the shock occurs. 3 The shock wave occurs at the exit plane. Properties The constant-pressure specific heat and the specific heat ratio of air are cp ! 1.005 kJ/kg · K and k ! 1.4. The gas constant of air is 0.287 kJ/kg % K. Analysis (a) The fluid properties at the exit of the nozzle just before the shock (denoted by subscript 1) are those evaluated in Example 12–7 at the nozzle exit to be P01 ! 1.0 MPa
P1 ! 0.1278 MPa
T1 ! 444.5 K
r 1 ! 1.002 kg/m3
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639 CHAPTER 12
The fluid properties after the shock (denoted by subscript 2) are related to those before the shock through the functions listed in Table A–14. For Ma1 ! 2.0, we read
Ma2 ! 0.5774
P02 ! 0.7209 P01
P2 ! 4.5000 P1
T2 ! 1.6875 T1
r2 ! 2.6667 r1
Then the stagnation pressure P02, static pressure P2, static temperature T2, and static density r2 after the shock are
P02 ! 0.7209P01 ! (0.7209)(1.0 MPa) ! 0.721 MPa P2 ! 4.5000P1 ! (4.5000)(0.1278 MPa) ! 0.575 MPa T2 ! 1.6875T1 ! (1.6875)(444.5 K) ! 750 K r 2 ! 2.6667r 1 ! (2.6667)(1.002 kg/m3) ! 2.67 kg/m3 (b) The entropy change across the shock is
s2 $ s1 ! cr ln
T2 P2 $ R ln T1 P1
! (1.005 kJ/kg % K) ln (1.6875) $ (0.287 kJ/kg % K) ln (4.5000) ! 0.0942 kJ/kg % K Thus, the entropy of the air increases as it experiences a normal shock, which is highly irreversible. (c) The air velocity after the shock can be determined from V2 ! Ma2c2, where c2 is the speed of sound at the exit conditions after the shock:
V2 ! Ma2c2 ! Ma2 2kRT2 ! (0.5774)
B
(1.4)(0.287 kJ/kg % K)(750.1 K)a
! 317 m/s
1000 m2/s2 b 1 kJ/kg
(d) The mass flow rate through a converging–diverging nozzle with sonic conditions at the throat is not affected by the presence of shock waves in the nozzle. Therefore, the mass flow rate in this case is the same as that determined in Example 12–7:
# m ! 2.86 kg/s Discussion This result can easily be verified by using property values at the nozzle exit after the shock at all Mach numbers significantly greater than unity.
Example 12–9 illustrates that the stagnation pressure and velocity decrease while the static pressure, temperature, density, and entropy increase across the shock. The rise in the temperature of the fluid downstream of a shock wave is of major concern to the aerospace engineer because it creates heat transfer problems on the leading edges of wings and nose cones of space reentry vehicles and the recently proposed hypersonic space planes. Overheating, in fact, led to the tragic loss of the space shuttle Columbia in February of 2003 as it was reentering earth’s atmosphere.
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640 FLUID MECHANICS
FIGURE 12–36 Schlieren image of a small model of the space shuttle Orbiter being tested at Mach 3 in the supersonic wind tunnel of the Penn State Gas Dynamics Lab. Several oblique shocks are seen in the air surrounding the spacecraft. Photo by G. S. Settles, Penn State University. Used by permission.
Oblique Shocks
Oblique shock Ma1 Ma 2 Ma1
u
b d
FIGURE 12–37 An oblique shock of shock angle b formed by a slender, two-dimensional wedge of half-angle d. The flow is turned by deflection angle u downstream of the shock, and the Mach number decreases.
Not all shock waves are normal shocks (perpendicular to the flow direction). For example, when the space shuttle travels at supersonic speeds through the atmosphere, it produces a complicated shock pattern consisting of inclined shock waves called oblique shocks (Fig. 12–36). As you can see, some portions of an oblique shock are curved, while other portions are straight. First, we consider straight oblique shocks, like that produced when a uniform supersonic flow (Ma1 ( 1) impinges on a slender, two-dimensional wedge of half-angle d (Fig. 12–37). Since information about the wedge cannot travel upstream in a supersonic flow, the fluid “knows” nothing about the wedge until it hits the nose. At that point, since the fluid cannot flow through the wedge, it turns suddenly through an angle called the turning angle or deflection angle u. The result is a straight oblique shock wave, aligned at shock angle or wave angle b, measured relative to the oncoming flow (Fig. 12–38). To conserve mass, b must obviously be greater than d. Since the Reynolds number of supersonic flows is typically large, the boundary layer growing along the wedge is very thin, and we ignore its effects. The flow therefore turns by the same angle as the wedge; namely, deflection angle u is equal to wedge half-angle d. If we take into account the displacement thickness effect of the boundary layer (Chap. 10), the deflection angle u of the oblique shock turns out to be slightly greater than wedge half-angle d. Like normal shocks, the Mach number decreases across an oblique shock, and oblique shocks are possible only if the upstream flow is supersonic. However, unlike normal shocks, in which the downstream Mach number is always subsonic, Ma2 downstream of an oblique shock can be subsonic, sonic, or supersonic, depending on the upstream Mach number Ma1 and the turning angle.
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641 CHAPTER 12
We analyze a straight oblique shock in Fig. 12–38 by decomposing the velocity vectors upstream and downstream of the shock into normal and tangential components, and considering a small control volume around the shock. Upstream of the shock, all fluid properties (velocity, density, pressure, etc.) along the lower left face of the control volume are identical to those along the upper right face. The same is true downstream of the shock. Therefore, the mass flow rates entering and leaving those two faces cancel each other out, and conservation of mass reduces to r 1V1, n A ! r 2V2, n A → r 1V1, n ! r 2V2, n
P1
V1, t V1, n
P2 →
V2
→
V1
u
Control volume
V2, t V2, n
b
(12–41)
where A is the area of the control surface that is parallel to the shock. Since A is the same on either side of the shock, it has dropped out of Eq. 12–41. As you might expect, the tangential component of velocity (parallel to the oblique shock) does not change across the shock, i.e., V1, t ! V2, t. This is easily proven by applying the tangential momentum equation to the control volume. When we apply conservation of momentum in the direction normal to the oblique shock, the only forces are pressure forces, and we get P1 A $ P2 A ! rV2, n AV2, n $ rV1, n AV1, n → P1 $ P2 ! r 2V 22, n $ r 1V 21, n
Oblique shock
FIGURE 12–38 Velocity vectors through an oblique shock of shock angle b and deflection angle u.
(12–42)
Finally, since there is no work done by the control volume and no heat transfer into or out of the control volume, stagnation enthalpy does not change across an oblique shock, and conservation of energy yields 1 1 1 1 h01 ! h02 ! h0 → h1 " V 21, n " V 21, t ! h2 " V 22, n " V 22, t 2 2 2 2
But since V1, t ! V2, t , this equation reduces to 1 1 h 1 " V 21, n ! h2 " V 22, n 2 2
(12–43)
Careful comparison reveals that the equations for conservation of mass, momentum, and energy (Eqs. 12–41 through 12–43) across an oblique shock are identical to those across a normal shock, except that they are written in terms of the normal velocity component only. Therefore, the normal shock relations derived previously apply to oblique shocks as well, but must be written in terms of Mach numbers Ma1, n and Ma2, n normal to the oblique shock. This is most easily visualized by rotating the velocity vectors in Fig. 12–38 by angle p/2 $ b, so that the oblique shock appears to be vertical (Fig. 12–39). Trigonometry yields Ma1, n ! Ma1 sin b
and
Ma2, n ! Ma2 sin(b $ u)
(12–44)
where Ma1, n ! V1, n/c1 and Ma2, n ! V2, n/c2. From the point of view shown in Fig. 12–39, we see what looks like a normal shock, but with some superposed tangential flow “coming along for the ride.” Thus, All the equations, shock tables, etc., for normal shocks apply to oblique shocks as well, provided that we use only the normal components of the Mach number.
In fact, you may think of normal shocks as special oblique shocks in which shock angle b ! p/2, or 90°. We recognize immediately that an oblique shock can exist only if Ma1, n ( 1 and Ma2, n ' 1. The normal
Ma 2, n ' 1
Ma 1, n ( 1
b$u
u
Oblique shock →
V2 V2, t V1, n V2, n
V1, t b
→
V1 P1
P2
FIGURE 12–39 The same velocity vectors of Fig. 12–38, but rotated by angle p/2 $ b, so that the oblique shock is vertical. Normal Mach numbers Man, 1 and Man, 2 are also defined.
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642 FLUID MECHANICS →
h01 ! h02
(k $ 1)Ma 21, n " 2
Ma 2, n ! P2 P1 r2 r1 T2 T1 P02 P01
!
!
B 2k Ma 21, n $ k " 1 2k Ma 21, n $ k " 1
V1, n V2, n
k"1 !
(k " 1)Ma 21, n 2 " (k $ 1)Ma 21, n
! [2 " (k $ 1)Ma21, n ]
!c
(k "
T01 ! T02
1)Ma 21, n
2"(k $ 1)Ma 21, n
2k Ma 21, n$ k " 1 (k " 1)2Ma 21, n
k/(k$1)
d
c
(k " 1)
1/(k$1)
d
2k Ma 21, n $ k " 1
FIGURE 12–40 Relationships across an oblique shock for an ideal gas in terms of the normal component of upstream Mach number Ma1, n.
shock equations appropriate for oblique shocks in an ideal gas are summarized in Fig. 12–40 in terms of Ma1, n. For known shock angle b and known upstream Mach number Ma1, we use the first part of Eq. 12–44 to calculate Ma1, n, and then use the normal shock tables (or their corresponding equations) to obtain Ma2, n. If we also knew the deflection angle u, we could calculate Ma2 from the second part of Eq. 12–44. But, in a typical application, we know either b or u, but not both. Fortunately, a bit more algebra provides us with a relationship between u, b, and Ma1. We begin by noting that tan b ! V1, n /V1, t and tan(b $ u) ! V2, n /V2, t (Fig. 12–39). But since V1, t ! V2, t, we combine these two expressions to yield V2, n V1, n
!
tan(b $ u) 2 " (k $ 1)Ma21, n 2 " (k $ 1)Ma21 sin2 b ! ! tan b (k " 1)Ma21, n (k " 1)Ma21 sin2 b
(12–45)
where we have also used Eq. 12–44 and the fourth equation of Fig. 12–40. We apply trigonometric identities for cos 2b and tan(b $ u), namely, cos 2b ! cos2 b $ sin2 b
and
tan(b $ u) !
tan b $ tan u 1 " tan b tan u
After some algebra, Eq. 12–45 reduces to The u-b-Ma relationship:
tan u !
2 cot b(Ma21 sin2 b $ 1) Ma21(k " cos 2b) " 2
(12–46)
Equation 12–46 provides deflection angle u as a unique function of shock angle b, specific heat ratio k, and upstream Mach number Ma1. For air (k ! 1.4), we plot u versus b for several values of Ma1 in Fig. 12–41. We note that this plot is often presented with the axes reversed (b versus u) in compressible flow textbooks, since, physically, shock angle b is determined by deflection angle u. Much can be learned by studying Fig. 12–41, and we list some observations here: • Figure 12–41 displays the full range of possible shock waves at a given free-stream Mach number, from the weakest to the strongest. For any value of Mach number Ma1 greater than 1, the possible values of u range from u ! 0° at some value of b between 0 and 90°, to a maximum value u ! umax at an intermediate value of b, and then back to u ! 0° at b ! 90°. Straight oblique shocks for u or b outside of this range cannot and do not exist. At Ma1 ! 1.5, for example, straight oblique shocks cannot exist in air with shock angle b less than about 42°, nor with deflection angle u greater than about 12°. If the wedge half-angle is greater than umax, the shock becomes curved and detaches from the nose of the wedge, forming what is called a detached oblique shock or a bow wave (Fig. 12–42). The shock angle b of the detached shock is 90° at the nose, but b decreases as the shock curves downstream. Detached shocks are much more complicated than simple straight oblique shocks to analyze. In fact, no simple solutions exist, and prediction of detached shocks requires computational methods (Chap. 15). • Similar oblique shock behavior is observed in axisymmetric flow over cones, as in Fig. 12–43, although the u-b-Ma relationship for axisymmetric flows differs from that of Eq. 12–46.
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643 CHAPTER 12 50 Ma2 ! 1
u ! umax
u, degrees
40
30 Ma2 ( 1
Ma1 → *
Ma2 ' 1
20 Weak
Strong
10 10 0
5
3
2
1.5 1.2
0
10
20
30
50 40 b, degrees
60
70
80
90
• When supersonic flow impinges on a blunt body—a body without a sharply pointed nose, the wedge half-angle d at the nose is 90°, and an attached oblique shock cannot exist, regardless of Mach number. In fact, a detached oblique shock occurs in front of all such blunt-nosed bodies, whether two-dimensional, axisymmetric, or fully three-dimensional. For example, a detached oblique shock is seen in front of the space shuttle model in Fig. 12–36 and in front of a sphere in Fig. 12–44. • While u is a unique function of Ma1 and b for a given value of k, there are two possible values of b for u ' umax. The dashed black line in Fig. 12–41 passes through the locus of umax values, dividing the shocks into weak oblique shocks (the smaller value of b) and strong oblique shocks (the larger value of b). At a given value of u, the weak shock is more common and is “preferred” by the flow unless the downstream pressure conditions are high enough for the formation of a strong shock. • For a given upstream Mach number Ma1, there is a unique value of u for which the downstream Mach number Ma2 is exactly 1. The dashed gray line in Fig. 12–41 passes through the locus of values where Ma2 ! 1. To the left of this line, Ma2 ( 1, and to the right of this line, Ma2 ' 1. Downstream sonic conditions occur on the weak shock side of the plot, with u very close to umax. Thus, the flow downstream of a strong oblique shock is always subsonic (Ma2 ' 1). The flow downstream of a weak oblique shock remains supersonic, except for a narrow range of u just below umax, where it is subsonic, although it is still called a weak oblique shock. • As the upstream Mach number approaches infinity, straight oblique shocks become possible for any b between 0 and 90°, but the maximum possible turning angle for k ! 1.4 (air) is umax ! 45.6°, which occurs at b ! 67.8°. Straight oblique shocks with turning angles above this value of umax are not possible, regardless of the Mach number. • For a given value of upstream Mach number, there are two shock angles where there is no turning of the flow (u ! 0°): the strong case, b ! 90°,
FIGURE 12–41 The dependence of straight oblique shock deflection angle u on shock angle b for several values of upstream Mach number Ma1. Calculations are for an ideal gas with k ! 1.4. The dashed black line connects points of maximum deflection angle (u ! umax). Weak oblique shocks are to the left of this line, while strong oblique shocks are to the right of this line. The dashed gray line connects points where the downstream Mach number is sonic (Ma2 ! 1). Supersonic downstream flow (Ma2 ( 1) is to the left of this line, while subsonic downstream flow (Ma2 ' 1) is to the right of this line.
Detached oblique shock Ma1
d ( umax
FIGURE 12–42 A detached oblique shock occurs upstream of a two-dimensional wedge of half-angle d when d is greater than the maximum possible deflection angle u. A shock of this kind is called a bow wave because of its resemblance to the water wave that forms at the bow of a ship.
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644 FLUID MECHANICS
FIGURE 12–43 Still frames from schlieren videography illustrating the detachment of an oblique shock from a cone with increasing cone half-angle d in air at Mach 3. At (a) d ! 20° and (b) d ! 40°, the oblique shock remains attached, but by (c) d ! 60°, the oblique shock has detached, forming a bow wave.
(a)
(b)
(c)
Ma1 d
Photos by G. S. Settles, Penn State University. Used by permission.
corresponds to a normal shock, and the weak case, b ! bmin, represents the weakest possible oblique shock at that Mach number, which is called a Mach wave. Mach waves are caused, for example, by very small nonuniformities on the walls of a supersonic wind tunnel (several can be seen in Figs. 12–36 and 12–43). Mach waves have no effect on the flow, since the shock is vanishingly weak. In fact, in the limit, Mach waves are isentropic. The shock angle for Mach waves is a unique function of the Mach number and is given the symbol m, not to be confused with the coefficient of viscosity. Angle m is called the Mach angle and is found by setting u equal to zero in Eq. 12–46, solving for b ! m, and taking the smaller root. We get Mach angle:
m ! sin$1(1/Ma1)
(12–47)
Since the specific heat ratio appears only in the denominator of Eq. 12–46, m is independent of k. Thus, we can estimate the Mach number of any supersonic flow simply by measuring the Mach angle and applying Eq. 12–47.
Prandtl–Meyer Expansion Waves
FIGURE 12–44 Shadowgram of a 12-in diameter sphere in free flight through air at Ma ! 1.53. The flow is subsonic behind the part of the bow wave that is ahead of the sphere and over its surface back to about 45°. At about 90° the laminar boundary layer separates through an oblique shock wave and quickly becomes turbulent. The fluctuating wake generates a system of weak disturbances that merge into the second “recompression” shock wave. Photo by A. C. Charters, as found in Van Dyke (1982).
We now address situations where supersonic flow is turned in the opposite direction, such as in the upper portion of a two-dimensional wedge at an angle of attack greater than its half-angle d (Fig. 12–45). We refer to this type of flow as an expanding flow, whereas a flow that produces an oblique shock may be called a compressing flow. As previously, the flow changes direction to conserve mass. However, unlike a compressing flow, an expanding flow does not result in a shock wave. Rather, a continuous expanding region called an expansion fan appears, composed of an infinite number of Mach waves called Prandtl–Meyer expansion waves. In other words, the flow does not turn suddenly, as through a shock, but gradually—each successive Mach wave turns the flow by an infinitesimal amount. Since each individual expansion wave is isentropic, the flow across the entire expansion fan is also isentropic. The Mach number downstream of the expansion increases (Ma2 ( Ma1), while pressure, density, and temperature decrease, just as they do in the supersonic (expanding) portion of a converging–diverging nozzle. Prandtl–Meyer expansion waves are inclined at the local Mach angle m, as sketched in Fig. 12–45. The Mach angle of the first expansion wave is
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645 CHAPTER 12
easily determined as m1 ! sin$1(1/Ma1). Similarly, m2 ! sin$1(1/Ma2), where we must be careful to measure the angle relative to the new direction of flow downstream of the expansion, namely, parallel to the upper wall of the wedge in Fig. 12–45 if we neglect the influence of the boundary layer along the wall. But how do we determine Ma2? It turns out that the turning angle u across the expansion fan can be calculated by integration, making use of the isentropic flow relationships. For an ideal gas, the result is (Anderson, 2003), Turning angle across an expansion fan:
u ! n(Ma2) $ n(Ma1)
(12–48)
where n(Ma) is an angle called the Prandtl–Meyer function (not to be confused with the kinematic viscosity), n(Ma) !
k " 1 $1 k$1 (Ma2 $ 1)b $ tan$1 a 2Ma2 $ 1b tan a Bk $ 1 Bk " 1
(12–49)
Note that n(Ma) is an angle, and can be calculated in either degrees or radians. Physically, n(Ma) is the angle through which the flow must expand, starting with n ! 0 at Ma ! 1, in order to reach a supersonic Mach number, Ma ( 1. To find Ma2 for known values of Ma1, k, and u, we calculate n(Ma1) from Eq. 12–49, n(Ma2) from Eq. 12–48, and then Ma2 from Eq. 12–49, noting that the last step involves solving an implicit equation for Ma2. Since there is no heat transfer or work, and the flow can be approximated as isentropic through the expansion, T0 and P0 remain constant, and we use the isentropic flow relations derived previously to calculate other flow properties downstream of the expansion, such as T2, r2, and P2. Prandtl–Meyer expansion fans also occur in axisymmetric supersonic flows, as in the corners and trailing edges of a cone-cylinder (Fig. 12–46). Some very complex and, to some of us, beautiful interactions involving both shock waves and expansion waves occur in the supersonic jet produced by an “overexpanded” nozzle, as in Fig. 12–47. Analysis of such flows is beyond the scope of the present text; interested readers are referred to compressible flow textbooks such as Thompson (1972), Leipmann and Roshko (2001), and Anderson (2003).
Expansion waves
m1 Ma1 ( 1
Ma 2 m2 u d
Oblique shock
FIGURE 12–45 An expansion fan in the upper portion of the flow formed by a two-dimensional wedge at the angle of attack in a supersonic flow. The flow is turned by angle u, and the Mach number increases across the expansion fan. Mach angles upstream and downstream of the expansion fan are indicated. Only three expansion waves are shown for simplicity, but in fact, there are an infinite number of them. (An oblique shock is present in the bottom portion of this flow.)
FIGURE 12–46 A cone-cylinder of 12.5° half-angle in a Mach number 1.84 flow. The boundary layer becomes turbulent shortly downstream of the nose, generating Mach waves that are visible in this shadowgraph. Expansion waves are seen at the corners and at the trailing edge of the cone. Photo by A. C. Charters, as found in Van Dyke (1982).
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646 FLUID MECHANICS
FIGURE 12–47 The complex interactions between shock waves and expansion waves in an “overexpanded” supersonic jet. The flow is visualized by a schlieren-like differential interferogram. Photo by Julio Srulijes. Reproduced by courtesy of the French-German Research Institute of SaintLouis, ISL. Used by permission.
EXAMPLE 12–10
Estimation of the Mach Number from Mach Lines
Estimate the Mach number of the free-stream flow upstream of the space shuttle in Fig. 12–36 from the figure alone. Compare with the known value of Mach number provided in the figure caption.
SOLUTION We are to estimate the Mach number from a figure and compare it to the known value. Analysis Using a protractor, we measure the angle of the Mach lines in the free-stream flow: m ! 19°. The Mach number is obtained from Eq. 12–47,
Weak shock
Ma1
bweak d ! 10°
m ! sin$1 a
1 1 b → Ma1 ! → Ma1 ! 3.07 Ma1 sin 19)
Our estimated Mach number agrees with the experimental value of 3.0 , 0.1. Discussion The result is independent of the fluid properties.
(a) Strong shock
EXAMPLE 12–11
Ma1 bstrong d ! 10°
(b)
FIGURE 12–48 Two possible oblique shock angles, (a) bweak and (b) bstrong, formed by a two-dimensional wedge of half-angle d ! 10°.
Oblique Shock Calculations
Supersonic air at Ma1 ! 2.0 and 75.0 kPa impinges on a two-dimensional wedge of half-angle d ! 10° (Fig. 12–48). Calculate the two possible oblique shock angles, bweak and bstrong, that could be formed by this wedge. For each case, calculate the pressure and Mach number downstream of the oblique shock, compare, and discuss.
SOLUTION We are to calculate the shock angle, Mach number, and pressure downstream of the weak and strong oblique shock formed by a twodimensional wedge. Assumptions 1 The flow is steady. 2 The boundary layer on the wedge is very thin. Properties The fluid is air with k ! 1.4.
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647 CHAPTER 12
Analysis Because of assumption 2, we approximate the oblique shock deflection angle as equal to the wedge half-angle, i.e., u ! d ! 10°. With Ma1 ! 2.0 and u ! 10°, we solve Eq. 12–46 for the two possible values of oblique shock angle b: Bweak ! 39.3° and Bstrong ! 83.7°. From these values, we use the first part of Eq. 12–44 to calculate upstream normal Mach number Ma1, n,
Weak shock:
Ma1, n ! Ma1 sin b → Ma1, n ! 2.0 sin 39.3) ! 1.267
and
Strong shock:
Ma1, n ! Ma1 sin b → Ma1, n ! 2.0 sin 83.7) ! 1.988
We substitute these values of Ma1, n into the second equation of Fig. 12–40 to calculate the downstream normal Mach number Ma2, n. For the weak shock, Ma2, n ! 0.8032, and for the strong shock, Ma2, n ! 0.5794. We also calculate the downstream pressure for each case, using the third equation of Fig. 12–40, which gives
Weak shock: P2 2k Ma21, n $ k " 1 2(1.4)(1.267)2 $ 1.4 " 1 ! → P2 ! (75.0 kPa) ! 128 kPa P1 k"1 1.4 " 1 and
Strong shock: P2 2k Ma21, n $ k " 1 2(1.4)(1.988)2 $ 1.4 " 1 ! → P2 ! (75.0 kPa) ! 333 kPa P1 k"1 1.4 " 1 Finally, we use the second part of Eq. 12–44 to calculate the downstream Mach number,
Weak shock:
Ma2 !
Ma2, n sin(b $ u)
!
0.8032 ! 1.64 sin(39.3) $ 10))
and
Strong shock:
Ma2 !
Ma2, n sin(b $ u)
!
0.5794 ! 0.604 sin(83.7) $ 10))
The changes in Mach number and pressure across the strong shock are much greater than the changes across the weak shock, as expected. Discussion Since Eq. 12–46 is implicit in b, we solve it by an iterative approach or with an equation solver such as EES. For both the weak and strong oblique shock cases, Ma1, n is supersonic and Ma2, n is subsonic. However, Ma2 is supersonic across the weak oblique shock, but subsonic across the strong oblique shock. We could also use the normal shock tables in place of the equations, but with loss of precision.
Ma1 ! 2.0
u Ma 2
EXAMPLE 12–12
Prandtl–Meyer Expansion Wave Calculations
Supersonic air at Ma1 ! 2.0 and 230 kPa flows parallel to a flat wall that suddenly expands by d ! 10° (Fig. 12–49). Ignoring any effects caused by the boundary layer along the wall, calculate downstream Mach number Ma2 and pressure P2.
d ! 10°
FIGURE 12–49 An expansion fan caused by the sudden expansion of a wall with d ! 10°.
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648 FLUID MECHANICS
SOLUTION We are to calculate the Mach number and pressure downstream of a sudden expansion along a wall. Assumptions 1 The flow is steady. 2 The boundary layer on the wall is very thin. Properties The fluid is air with k ! 1.4. Analysis Because of assumption 2, we approximate the total deflection angle as equal to the wall expansion angle, i.e., u ! d ! 10°. With Ma1 ! 2.0, we solve Eq. 12–49 for the upstream Prandtl–Meyer function, n(Ma) !
!
k$1 k " 1 $1 tan a (Ma2 $ 1)b $ tan$1 a 2Ma2 $ 1b Bk " 1 Bk $ 1
1.4$1 1.4"1 $1 tan a (2.0 2 $1)b $ tan$1 a22.0 2 $1b ! 26.38° B 1.4"1 B 1.4$1
Next, we use Eq. 12–48 to calculate the downstream Prandtl–Meyer function,
u ! n(Ma2) $ n(Ma1) → n(Ma2) ! u " n(Ma1) ! 10) " 26.38) ! 36.38) Ma2 is found by solving Eq. 12–49, which is implicit—an equation solver is helpful. We get Ma2 ! 2.385. There are also compressible flow calculators on the Internet that solve these implicit equations, along with both normal and oblique shock equations; e.g., see www.aoe.vt.edu/~devenpor/aoe3114/ calc.html. We use the isentropic relations to calculate the downstream pressure, $k/(k$1) k$1 b Ma22d P2/P0 2 P2 ! P ! (230 kPa) ! 126 kPa $k/(k$1) P1/P0 1 k$1 c1 " a b Ma21d 2
c1 " a
Since this is an expansion, Mach number increases and pressure decreases, as expected. Discussion We could also solve for downstream temperature, density, etc., using the appropriate isentropic relations.
Fuel nozzles or spray bars
Air inlet
Flame holders
FIGURE 12–50 Many practical compressible flow problems involve combustion, which may be modeled as heat gain through the duct wall.
12–6
■
DUCT FLOW WITH HEAT TRANSFER AND NEGLIGIBLE FRICTION (RAYLEIGH FLOW)
So far we have limited our consideration mostly to isentropic flow, also called reversible adiabatic flow since it involves no heat transfer and no irreversibilities such as friction. Many compressible flow problems encountered in practice involve chemical reactions such as combustion, nuclear reactions, evaporation, and condensation as well as heat gain or heat loss through the duct wall. Such problems are difficult to analyze exactly since they may involve significant changes in chemical composition during flow, and the conversion of latent, chemical, and nuclear energies to thermal energy (Fig. 12–50). The essential features of such complex flows can still be captured by a simple analysis by modeling the generation or absorption of thermal energy as heat transfer through the duct wall at the same rate and disregarding any
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649 CHAPTER 12
changes in chemical composition. This simplified problem is still too complicated for an elementary treatment of the topic since the flow may involve friction, variations in duct area, and multidimensional effects. In this section, we limit our consideration to one-dimensional flow in a duct of constant cross-sectional area with negligible frictional effects. Consider steady one-dimensional flow of an ideal gas with constant specific heats through a constant-area duct with heat transfer, but with negligible friction. Such flows are referred to as Rayleigh flows after Lord Rayleigh (1842–1919). The conservation of mass, momentum, and energy equations for the control volume shown in Fig. 12–51 can be written as follows: Continuity equation Noting that the duct cross-sectional area A is . . constant, the relation m1 ! m2 or r1A1V1 ! r2 A2V2 reduces to r 1V1 ! r 2V2
(12–50)
x-Momentum equation Noting that the frictional effects are negligible and thus there are no shear forces, and assuming there are no external → #→ # → and body forces, the momentum equation a F ! a bm V $ a bmV out
in
in the flow (or x-) direction becomes a balance between static pressure forces and momentum transfer. Noting that the flows are high speed and turbulent, the momentum flux correction factor is approximately 1 (b ! 1) and thus can be neglected. Then, # # P1A 1 $ P2A 2 ! mV2 $ mV1 → P1 $ P2 ! (r 2V2)V2 $ (r 1V1)V1
or P1 " r 1V 21 ! P2 " r 2V 22
(12–51)
Energy equation The control volume involves no shear, shaft, or other forms of work, and the .potential energy change is negligible. If the rate of. heat transfer is Q and the heat transfer . per. unit mass of fluid is . q ! Q/m, the steady-flow energy balance E in ! E out becomes # V 21 V 22 # # Q " m ah 1 " b ! m ah2 " b 2 2
→ q " h1 "
V 21 V 22 ! h2 " (12–52) 2 2
For an ideal gas with constant specific heats, -h ! cp -T, and thus q ! cp(T2 $ T1) "
V 22 $ V 21 2
(12–53)
or q ! h02 $ h01 ! cp(T02 $ T01)
(12–54)
Therefore, the stagnation enthalpy h0 and stagnation temperature T0 change during Rayleigh flow (both increase when heat is transferred to the fluid and thus q is positive, and both decrease when heat is transferred from the fluid and thus q is negative). Entropy change In the absence of any irreversibilities such as friction, the entropy of a system changes by heat transfer only: it increases with heat gain, and decreases with heat loss. Entropy is a property and thus
.
Q P1, T 1, r 1
P2 , T 2 , r 2
V1
V2 Control volume
FIGURE 12–51 Control volume for flow in a constantarea duct with heat transfer and negligible friction.
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650 FLUID MECHANICS
a state function, and the entropy change of an ideal gas with constant specific heats during a change of state from 1 to 2 is given by s2 $ s1 ! cp ln
T2 P2 $ R ln T1 P1
(12–55)
The entropy of a fluid may increase or decrease during Rayleigh flow, depending on the direction of heat transfer. Equation of state Noting that P ! rRT, the properties P, r, and T of an ideal gas at states 1 and 2 are related to each other by P2 P1 ! r 1T1 r 2T2
T
Mab = 1/ k
Tmax Cooling (Ma S 0)
b
Ma < 1
Heating (Ma S 1) Ma > 1 Heating (Ma S 1)
a Maa = 1
smax
Cooling (Ma S *) s
FIGURE 12–52 T-s diagram for flow in a constant-area duct with heat transfer and negligible friction (Rayleigh flow).
(12–56)
Consider a gas with known properties R, k, and cp. For a specified inlet state 1, the inlet properties P1, T1, r1, V1, and s1 are known. The five exit properties P2, T2, r2, V2, and s2 can be determined from the five equations 12–50, 12–51, 12–53, 12–55, and 12–56 for any specified value of heat transfer q. When the velocity and temperature are known, the Mach number can be determined from Ma ! V/c ! V/ 1kRT . Obviously there is an infinite number of possible downstream states 2 corresponding to a given upstream state 1. A practical way of determining these downstream states is to assume various values of T2, and calculate all other properties as well as the heat transfer q for each assumed T2 from the Eqs. 12–50 through 12–56. Plotting the results on a T-s diagram gives a curve passing through the specified inlet state, as shown in Fig. 12–52. The plot of Rayleigh flow on a T-s diagram is called the Rayleigh line, and several important observations can be made from this plot and the results of the calculations: 1. All the states that satisfy the conservation of mass, momentum, and energy equations as well as the property relations are on the Rayleigh line. Therefore, for a given initial state, the fluid cannot exist at any downstream state outside the Rayleigh line on a T-s diagram. In fact, the Rayleigh line is the locus of all physically attainable downstream states corresponding to an initial state. 2. Entropy increases with heat gain, and thus we proceed to the right on the Rayleigh line as heat is transferred to the fluid. The Mach number is Ma ! 1 at point a, which is the point of maximum entropy (see Example 12–13 for proof). The states on the upper arm of the Rayleigh line above point a are subsonic, and the states on the lower arm below point a are supersonic. Therefore, a process proceeds to the right on the Rayleigh line with heat addition and to the left with heat rejection regardless of the initial value of the Mach number. 3. Heating increases the Mach number for subsonic flow, but decreases it for supersonic flow. The flow Mach number approaches Ma ! 1 in both cases (from 0 in subsonic flow and from * in supersonic flow) during heating. 4. It is clear from the energy balance q ! cp(T02 $ T01) that heating increases the stagnation temperature T0 for both subsonic and supersonic flows, and cooling decreases it. (The maximum value of T0 occurs at Ma ! 1.) This is also the case for the static temperature T except for the
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651 CHAPTER 12
narrow Mach number range of 1/ 1k ' Ma ' 1 in subsonic flow (see Example 12–13). Both temperature and the Mach number increase with heating in subsonic flow, but T reaches a maximum Tmax at Ma ! 1/ 1k (which is 0.845 for air), and then decreases. It may seem peculiar that the temperature of a fluid drops as heat is transferred to it. But this is no more peculiar than the fluid velocity increasing in the diverging section of a converging–diverging nozzle. The cooling effect in this region is due to the large increase in the fluid velocity and the accompanying drop in temperature in accordance with the relation T0 ! T " V 2/2cp. Note also that heat rejection in the region 1/ 1k ' Ma ' 1 causes the fluid temperature to increase (Fig. 12–53). 5. The momentum equation P " KV ! constant, where K ! rV ! constant (from the continuity equation), reveals that velocity and static pressure have opposite trends. Therefore, static pressure decreases with heat gain in subsonic flow (since velocity and the Mach number increase), but increases with heat gain in supersonic flow (since velocity and the Mach number decrease). 6. The continuity equation rV ! constant indicates that density and velocity are inversely proportional. Therefore, density decreases with heat transfer to the fluid in subsonic flow (since velocity and the Mach number increase), but increases with heat gain in supersonic flow (since velocity and the Mach number decrease). 7. On the left half of Fig. 12–52, the lower arm of the Rayleigh line is steeper than the upper arm (in terms of s as a function of T), which indicates that the entropy change corresponding to a specified temperature change (and thus a given amount of heat transfer) is larger in supersonic flow.
Heating
T1
Subsonic flow
T2 ( T 1 or T2 ' T 1 T02 ( T 01
T01 Heating
T1
Supersonic flow
T2 ( T 1 T02 ( T 01
T01
FIGURE 12–53 During heating, fluid temperature always increases if the Rayleigh flow is supersonic, but the temperature may actually drop if the flow is subsonic.
The effects of heating and cooling on the properties of Rayleigh flow are listed in Table 12–3. Note that heating or cooling has opposite effects on most properties. Also, the stagnation pressure decreases during heating and increases during cooling regardless of whether the flow is subsonic or supersonic.
TA B L E 1 2 – 3 The effects of heating and cooling on the properties of Rayleigh flow Heating
Cooling
Property
Subsonic
Supersonic
Subsonic
Supersonic
Velocity, V Mach number, Ma Stagnation temperature, T0 Temperature, T
Increase Increase Increase Increase for Ma ' 1/k1/2 Decrease for Ma ( 1/k1/2 Decrease Decrease Decrease Increase
Decrease Decrease Increase Increase
Decrease Decrease Decrease Decrease for Ma ' 1/k1/2 Increase for Ma ( 1/k1/2 Increase Increase Increase Decrease
Increase Increase Decrease Decrease
Density, r Stagnation pressure, P0 Pressure, P Entropy, s
Increase Decrease Increase Increase
Decrease Increase Decrease Decrease
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652 FLUID MECHANICS
# ds $ ! 0 dT
T
Tmax
EXAMPLE 12–13
b
b a
Ma ' 1
# dT $ ! 0 ds
a
Ma ( 1
Consider the T-s diagram of Rayleigh flow, as shown in Fig. 12–54. Using the differential forms of the conservation equations and property relations, show that the Mach number is Maa ! 1 at the point of maximum entropy (point a), and Mab ! 1/ 1k at the point of maximum temperature (point b).
SOLUTION It is to be shown that Maa ! 1 at the point of maximum entropy
smax
s
FIGURE 12–54 The T-s diagram of Rayleigh flow considered in Example 12–13.
Extrema of Rayleigh Line
and Mab ! 1/ 1k at the point of maximum temperature on the Rayleigh line. Assumptions The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. Analysis The differential forms of the continuity (rV ! constant), momentum [rearranged as P " (rV)V ! constant], ideal gas (P ! rRT), and enthalpy change (-h ! cp -T) equations can be expressed as
→
rV ! constant
P " (rV)V ! constant P ! rRT
→
→
r dV " V dr ! 0 →
dP " (rV) dV ! 0
dP ! rR dT " RT dr
→
dr dV !$ r V →
dP ! $rV dV
dP dT dr ! " r P T
(1)
(2)
(3)
The differential form of the entropy change relation (Eq. 12–40) of an ideal gas with constant specific heats is
ds ! cp
dP dT $ R T P
(4)
Substituting Eq. 3 into Eq. 4 gives
ds ! cp since
dr dr R dT dT dr dT dT ! $ Ra " b ! (cp $ R) $R $R r r r T T T k$1 T cp $ R ! cv
→
kcv $ R ! cv
→
(5)
cv ! R/(k $ 1)
Dividing both sides of Eq. 5 by dT and combining with Eq. 1,
ds R R dV ! " dT T(k $ 1) V dT
(6)
Dividing Eq. 3 by dV and combining it with Eqs. 1 and 2 give, after rearranging,
dT T V ! $ dV V R
(7)
Substituting Eq. 7 into Eq. 6 and rearranging,
R(kRT $ V 2) ds R R ! " ! 2 dT T(k $ 1) T $ V /R T(k $ 1)(RT $ V 2)
(8)
Setting ds/dT ! 0 and solving the resulting equation R(kRT $ V 2) ! 0 for V give the velocity at point a to be
Va ! 2kRTa
and
Maa !
Va 2kRTa ! !1 ca 2kRT a
(9)
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653 CHAPTER 12
Therefore, sonic conditions exist at point a, and thus the Mach number is 1. Setting dT/ds ! (ds/dT)$1 ! 0 and solving the resulting equation T(k $ 1) # (RT $ V 2) ! 0 for velocity at point b give
Vb ! 2RTb
and
Mab !
Vb 2RTb 1 ! ! cb 2kRT 2k b
(10)
Therefore, the Mach number at point b is Mab ! 1/ 1k . For air, k ! 1.4 and thus Mab ! 0.845. Discussion Note that in Rayleigh flow, sonic conditions are reached as the entropy reaches its maximum value, and maximum temperature occurs during subsonic flow.
EXAMPLE 12–14
Effect of Heat Transfer on Flow Velocity
Starting with the differential form of the energy equation, show that the flow velocity increases with heat addition in subsonic Rayleigh flow, but decreases in supersonic Rayleigh flow.
SOLUTION It is to be shown that flow velocity increases with heat addition in subsonic Rayleigh flow and that the opposite occurs in supersonic flow. Assumptions 1 The assumptions associated with Rayleigh flow are valid. 2 There are no work interactions and potential energy changes are negligible. Analysis Consider heat transfer to the fluid in the differential amount of dq. The differential form of the energy equations can be expressed as dq ! dh0 ! dah "
V2 b ! cp dT " V dV 2
(1)
Dividing by cpT and factoring out dV/V give
dq dT V dV dV V dT (k $ 1)V 2 ! " ! a " b cpT T cpT V dV T kRT
(2)
where we also used cp ! kR/(k $ 1). Noting that Ma2 ! V 2/c2 ! V 2/kRT and using Eq. 7 for dT/dV from Example 12–13 give
dq dV V T V dV V2 ! a a $ b " (k $ 1)Ma2b ! a1 $ " k Ma2 $ Ma2b cpT V T V R V TR
dq (3)
Canceling the two middle terms in Eq. 3 since V 2/TR ! k Ma2 and rearranging give the desired relation,
dV dq 1 ! V cpT (1 $ Ma2)
V1
(4)
In subsonic flow, 1 $ Ma2 ( 0 and thus heat transfer and velocity change have the same sign. As a result, heating the fluid (dq ( 0) increases the flow velocity while cooling decreases it. In supersonic flow, however, 1 $ Ma2 ' 0 and heat transfer and velocity change have opposite signs. As a result, heating the fluid (dq ( 0) decreases the flow velocity while cooling increases it (Fig. 12–55). Discussion Note that heating the fluid has the opposite effect on flow velocity in subsonic and supersonic Rayleigh flows.
Subsonic flow
V2 ( V 1
dq
V1
Supersonic flow
V2 ' V 1
FIGURE 12–55 Heating increases the flow velocity in subsonic flow, but decreases it in supersonic flow.
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654 FLUID MECHANICS
Property Relations for Rayleigh Flow T0 (k " 1)Ma2[2 " (k $ 1)Ma2] ! T0* (1 " kMa Ma2)2 P0 P0*
!
1) (k $ 1)Ma2 k/(/(k$1) k " 1 2 "! a b 2 k " 1 1 " kMa Ma
Ma(1 " k) 2 Ma( T !a b T* 1 " kMa Ma2 P 1"k ! P* 1 " kMa Ma2
It is often desirable to express the variations in properties in terms of the Mach number Ma. Noting that Ma ! V/c ! V/ 1kRT and thus V ! Ma1kRT , rV 2 ! rkRTMa2 ! kPMa2
since P ! rRT. Substituting into the momentum equation (Eq. 12–51) gives P1 " kP1Ma21 ! P2 " kP2Ma22, which can be rearranged as P2 1 " kMa21 ! P1 1 " kMa22
V r* (1 " k)M )Ma2 V* ! r ! 1 " kMa Ma2
FIGURE 12–56 Summary of relations for Rayleigh flow.
(12–57)
(12–58)
Again utilizing V ! Ma 1kRT , the continuity equation r1V1 ! r2V2 can be expressed as r 1 V2 Ma2 2kRT2 Ma2 2T2 ! ! ! r 2 V1 Ma 2kRT Ma 2T 1
1
1
(12–59)
1
Then the ideal-gas relation (Eq. 12–56) becomes T2 P2 r 1 1 " kMa21 Ma2 2T2 ! !a ba b T1 P1 r 2 1 " kMa22 Ma1 2T1
(12–60)
Solving Eq. 12–60 for the temperature ratio T2/T1 gives T2 Ma2(1 " kMa21) 2 !a b T1 Ma1(1 " kMa22)
(12–61)
Substituting this relation into Eq. 12–59 gives the density or velocity ratio as r 2 V1 Ma21(1 " kMa22) ! ! r 1 V2 Ma22(1 " kMa21)
(12–62)
Flow properties at sonic conditions are usually easy to determine, and thus the critical state corresponding to Ma ! 1 serves as a convenient reference point in compressible flow. Taking state 2 to be the sonic state (Ma2 ! 1, and superscript * is used) and state 1 to be any state (no subscript), the property relations in Eqs. 12–58, 12–61, and 12–62 reduce to (Fig. 12–56) P 1"k ! P* 1 " kMa2
Ma(1 " k) 2 T b !a T* 1 " kMa2
and
r* (1 " k)Ma2 V ! (12–63) ! r V* 1 " kMa2
Similar relations can be obtained for dimensionless stagnation temperature and stagnation pressure as follows: T0 T0 T T* Ma(1 " k) 2 k$1 k $ 1 $1 2 (12–64) Ma b ! ! a1 " b a b a1 " 2 T*0 T T* T0* 2 2 1 " kMa
which simplifies to
(k " 1)Ma232 " (k $ 1)Ma24 T0 ! T0* (1 " kMa2)2
(12–65)
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655 CHAPTER 12
Also, k/(k$1) P0 P0 P P* k$1 1"k k $ 1 $k/(k$1) ! ! a1 " a b a1 " Ma2b b 2 * * P0 P P* P 0 2 2 1 " kMa
(12–66)
which simplifies to
P0 2 " (k $ 1)Ma2 k/(k$1) k"1 ! a b P0 * 1 " kMa2 k"1
(12–67)
The five relations in Eqs. 12–63, 12–65, and 12–67 enable us to calculate the dimensionless pressure, temperature, density, velocity, stagnation temperature, and stagnation pressure for Rayleigh flow of an ideal gas with a specified k for any given Mach number. Representative results are given in tabular and graphical form in Table A–15 for k ! 1.4.
Choked Rayleigh Flow
It is clear from the earlier discussions that subsonic Rayleigh flow in a duct may accelerate to sonic velocity (Ma ! 1) with heating. What happens if we continue to heat the fluid? Does the fluid continue to accelerate to supersonic velocities? An examination of the Rayleigh line indicates that the fluid at the critical state of Ma ! 1 cannot be accelerated to supersonic velocities by heating. Therefore, the flow is choked. This is analogous to not being able to accelerate a fluid to supersonic velocities in a converging nozzle by simply extending the converging flow section. If we keep heating the fluid, we will simply move the critical state further downstream and reduce the flow rate since fluid density at the critical state will now be lower. Therefore, for a given inlet state, the corresponding critical state fixes the maximum possible heat transfer for steady flow (Fig. 12–57). That is, qmax ! h*0 $ h01 ! cp(T *0 $ T01)
(12–68)
Further heat transfer causes choking and thus the inlet state to change (e.g., inlet velocity will decrease), and the flow no longer follows the same Rayleigh line. Cooling the subsonic Rayleigh flow reduces the velocity, and the Mach number approaches zero as the temperature approaches absolute zero. Note that the stagnation temperature T0 is maximum at the critical state of Ma ! 1. In supersonic Rayleigh flow, heating decreases the flow velocity. Further heating simply increases the temperature and moves the critical state farther downstream, resulting in a reduction in the mass flow rate of the fluid. It may seem like supersonic Rayleigh flow can be cooled indefinitely, but it turns out that there is a limit. Taking the limit of Eq. 12–65 as the Mach number approaches infinity gives LimMa→*
T0 1 !1$ 2 T *0 k
(12–69)
which yields T0/T*0 ! 0.49 for k ! 1.4. Therefore, if the critical stagnation temperature is 1000 K, air cannot be cooled below 490 K in Rayleigh flow. Physically this means that the flow velocity reaches infinity by the time the temperature reaches 490 K—a physical impossibility. When supersonic flow cannot be sustained, the flow undergoes a normal shock wave and becomes subsonic.
qmax T1 T01
Rayleigh flow
T2 ! T * * T02 ! T 01 Choked flow
FIGURE 12–57 For a given inlet state, the maximum possible heat transfer occurs when sonic conditions are reached at the exit state.
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656 FLUID MECHANICS
.
EXAMPLE 12–15
Q P1 ! 480 kPa T1 ! 550 K V1 ! 80 m/s
Combustor tube
P2, T 2, V 2
FIGURE 12–58 Schematic of the combustor tube analyzed in Example 12–15.
Rayleigh Flow in a Tubular Combustor
A combustion chamber consists of tubular combustors of 15-cm diameter. Compressed air enters the tubes at 550 K, 480 kPa, and 80 m/s (Fig. 12–58). Fuel with a heating value of 42,000 kJ/kg is injected into the air and is burned with an air–fuel mass ratio of 40. Approximating combustion as a heat transfer process to air, determine the temperature, pressure, velocity, and Mach number at the exit of the combustion chamber.
SOLUTION Fuel is burned in a tubular combustion chamber with compressed air. The exit temperature, pressure, velocity, and Mach number are to be determined. Assumptions 1 The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. 2 Combustion is complete, and it is treated as a heat addition process, with no change in the chemical composition of the flow. 3 The increase in mass flow rate due to fuel injection is disregarded. Properties We take the properties of air to be k ! 1.4, cp ! 1.005 kJ/kg · K, and R ! 0.287 kJ/kg · K. Analysis The inlet density and mass flow rate of air are r1 !
P1 480 kPa ! ! 3.041 kg/m3 RT1 (0.287 kJ/kg % K)(550 K)
# m air ! r 1A 1V1 ! (3.041 kg/m3) 3p(0.15 m)2/44(80 m/s) ! 4.299 kg/s
The mass flow rate of fuel and the rate of heat transfer are
# mair 4.299 kg/s # ! ! 0.1075 kg/s mfuel ! AF 40 # # Q ! mfuel HV ! (0.1075 kg/s)(42,000 kJ/kg) ! 4514 kW # Q 4514 kJ/s ! 1050 kJ/kg q! # ! m air 4.299 kg/s The stagnation temperature and Mach number at the inlet are
T01 ! T1 "
1 kJ/kg V 21 (80 m/s)2 ! 550 K " b ! 553.2 K a 2cp 2(1.005 kJ/kg % K) 1000 m2/s2
c1 ! 2kRT1 ! Ma1 !
B
(1.4)(0.287 kJ/kg % K)(550 K)a
V1 80 m/s ! ! 0.1702 c1 470.1 m/s
1000 m2/s2 b ! 470.1 m/s 1 kJ/kg
The exit stagnation temperature is, from the energy equation q ! cp(T02 $ T01),
T02 ! T01 "
1050 kJ/kg q ! 553.2 K " ! 1598 K cp 1.005 kJ/kg % K
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657 CHAPTER 12
The maximum value of stagnation temperature T *0 occurs at Ma ! 1, and its value can be determined from Table A–15 or from Eq. 12–65. At Ma1 ! 0.1702 we read T0/T *0 ! 0.1291. Therefore,
T*0 !
T01 553.2 K ! ! 4284 K 0.1291 0.1291
The stagnation temperature ratio at the exit state and the Mach number corresponding to it are, from Table A–15,
T02 1598 K ! ! 0.3730 → Ma2 ! 0.3142 T*0 4284 K The Rayleigh flow functions corresponding to the inlet and exit Mach numbers are (Table A–15):
Ma1 ! 0.1702:
T1 ! 0.1541 T*
P1 ! 2.3065 P*
V1 ! 0.0668 V*
Ma2 ! 0.3142:
T2 ! 0.4389 T*
P2 ! 2.1086 P*
V2 ! 0.2082 V*
Then the exit temperature, pressure, and velocity are determined to be
T2 T2/T* 0.4389 ! ! ! 2.848 → T2 ! 2.848T1 ! 2.848(550 K) ! 1566 K T1 T1/T* 0.1541 P2 P2/P* 2.1086 ! ! ! 0.9142 → P2 ! 0.9142P1 ! 0.9142(480 kPa) ! 439 kPa P1 P1/P* 2.3065 V2 V2/V* 0.2082 ! ! ! 3.117 → V2 ! 3.117V1 ! 3.117(80 m/s) ! 249 m/s V1 V1/V* 0.0668 Discussion Note that the temperature and velocity increase and pressure decreases during this subsonic Rayleigh flow with heating, as expected. This problem can also be solved using appropriate relations instead of tabulated values, which can likewise be coded for convenient computer solutions.
12–7
■
ADIABATIC DUCT FLOW WITH FRICTION (FANNO FLOW)
Wall friction associated with high-speed flow through short devices with large cross-sectional areas such as large nozzles is often negligible, and flow through such devices can be approximated as being frictionless. But wall friction is significant and should be considered when studying flows through long flow sections, such as long ducts, especially when the crosssectional area is small. In this section we consider compressible flow with significant wall friction but negligible heat transfer in ducts of constant cross-sectional area. Consider steady, one-dimensional, adiabatic flow of an ideal gas with constant specific heats through a constant-area duct with significant frictional
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658 FLUID MECHANICS x
Ffriction
P1, T 1, r 1
P2, T 2, r 2
V1
V2
Control volume
effects. Such flows are referred to as Fanno flows. The conservation of mass, momentum, and energy equations for the control volume shown in Fig. 12–59 can be written as follows: Continuity equation Noting that the duct cross-sectional area A is . . constant (and thus A1 ! A2 ! Ac), the relation m1 ! m2 or r1A1V1 ! r2 A2V2 reduces to
A1 ! A 2 ! A
FIGURE 12–59 Control volume for adiabatic flow in a constant-area duct with friction.
r 1V1 ! r 2V2 → rV ! constant
(12–70)
x-Momentum equation Denoting the friction force exerted on the fluid by the inner surface of the duct by Ffriction and assuming there are no other external and body forces, the momentum equation
→ # → # → a F ! a bmV $ a bmV in the flow direction can be expressed as out
in
Ffriction # # P1A $ P2 A $ Ffriction ! mV2 $ mV1 → P1 $ P2 $ A ! (r 2V2)V2 $ (r 1V1)V1
where even though there is friction at the walls, and the velocity profiles are not uniform, we approximate the momentum flux correction factor b as 1 for simplicity since the flow is turbulent. The equation can be rewritten as P1 " r 1V 21 ! P2 " r 2V 22 "
Ffriction A
(12–71)
Energy equation The control volume involves no heat or work interactions and the potential .energy. change is negligible. Then the steady-flow energy balance E in ! E out becomes h1 "
V 21 V 22 ! h2 " 2 2
→
h 01 ! h02 → h0 ! h "
V2 ! constant (12–72) 2
For an ideal gas with constant specific heats, -h ! cp -T and thus T1 "
V 21 V 22 ! T2 " 2cp 2cp
→
T01 ! T02 → T0 ! T "
V2 ! constant (12–73) 2cp
Therefore, the stagnation enthalpy h0 and stagnation temperature T0 remain constant during Fanno flow. Entropy change In the absence of any heat transfer, the entropy of a system can be changed only by irreversibilities such as friction, whose effect is always to increase entropy. Therefore, the entropy of the fluid must increase during Fanno flow. The entropy change in this case is equivalent to entropy increase or entropy generation, and for an ideal gas with constant specific heats it is expressed as s2 $ s1 ! cp ln
T2 P2 $ R ln ( 0 T1 P1
(12–74)
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659 CHAPTER 12
Equation of state Noting that P ! rRT, the properties P, r, and T of an ideal gas at states 1 and 2 are related to each other by P2 P1 ! r 1T1 r 2T2
T, K 1 500 Ma ' 1
(12–75)
Consider a gas with known properties R, k, and cp flowing in a duct of constant cross-sectional area A. For a specified inlet state 1, the inlet properties P1, T1, r1, V1, and s1 are known. The five exit properties P2, T2, r2, V2, and s2 can be determined from Eqs. 12–70 through 12–75 for any specified value of the friction force Ffriction. Knowing the velocity and temperature, we can also determine the Mach number at the inlet and the exit from the relation Ma ! V/c ! V/1kRT . Obviously there is an infinite number of possible downstream states 2 corresponding to a given upstream state 1. A practical way of determining these downstream states is to assume various values of T2, and calculate all other properties as well as the friction force for each assumed T2 from Eqs. 12–70 through 12–75. Plotting the results on a T-s diagram gives a curve passing through the specified inlet state, as shown in Fig. 12–60. The plot of Fanno flow on a T-s diagram is called the Fanno line, and several important observations can be made from this plot and the results of calculations: 1. All states that satisfy the conservation of mass, momentum, and energy equations as well as the property relations are on the Fanno line. Therefore, for a given inlet state, the fluid cannot exist at any downstream state outside the Fanno line on a T-s diagram. In fact, the Fanno line is the locus of all possible downstream states corresponding to an initial state. Note that if there were no friction, the flow properties would have remained constant along the duct during Fanno flow. 2. Friction causes entropy to increase, and thus a process always proceeds to the right along the Fanno line. At the point of maximum entropy, the Mach number is Ma ! 1. All states on the upper part of the Fanno line are subsonic, and all states on the lower part are supersonic. 3. Friction increases the Mach number for subsonic Fanno flow, but decreases it for supersonic Fanno flow. The Mach number approaches unity (Ma ! 1) in both cases. 4. The energy balance requires that stagnation temperature T0 ! T " V 2/2cp remain constant during Fanno flow. But the actual temperature may change. Velocity increases and thus temperature decreases during subsonic flow, but the opposite occurs during supersonic flow (Fig. 12–61). 5. The continuity equation rV ! constant indicates that density and velocity are inversely proportional. Therefore, the effect of friction is to decrease density in subsonic flow (since velocity and Mach number increase), but to increase it in supersonic flow (since velocity and Mach number decrease). The effects of friction on the properties of Fanno flow are listed in Table 12–4. Note that frictional effects on most properties in subsonic flow are opposite to those in supersonic flow. However, the effect of friction is to always decrease stagnation pressure, regardless of whether the flow is
400
Ma ! 1 and s ! smax
Ma ( 1 300
200
0
0.1
0.2
0.3 s, kJ/kg • K
FIGURE 12–60 T-s diagram for adiabatic frictional flow in a constant-area duct (Fanno flow). Numerical values are for air with k ! 1.4 and inlet conditions of T1 ! 500 K, P1 ! 600 kPa, V1 ! 80 m/s, and an assigned value of s1 ! 0.
Ffriction T1 Ma 1
Subsonic flow
T2 ' T1 Ma 2 ( Ma 1
Ffriction T1 Ma 1
Supersonic flow
T2 ( T1 Ma 2 ' Ma 1
FIGURE 12–61 Friction causes the Mach number to increase and the temperature to decrease in subsonic Fanno flow, but it does the opposite in supersonic Fanno flow.
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660 FLUID MECHANICS
TA B L E 1 2 – 4 The effects of friction on the properties of Fanno flow Property
Subsonic
Supersonic
Velocity, V Mach number, Ma Stagnation temperature, T0 Temperature, T Density, r Stagnation pressure, P0 Pressure, P Entropy, s
Increase Increase Constant Decrease Decrease Decrease Decrease Increase
Decrease Decrease Constant Increase Increase Decrease Increase Increase
subsonic or supersonic. But friction has no effect on stagnation temperature since friction simply causes the mechanical energy to be converted to an equivalent amount of thermal energy.
Property Relations for Fanno Flow
In compressible flow, it is convenient to express the variation of properties in terms of Mach number, and Fanno flow is no exception. However, Fanno flow involves the friction force, which is proportional to the square of the velocity even when the friction factor is constant. But in compressible flow, velocity varies significantly along the flow, and thus it is necessary to perform a differential analysis to account for the variation of the friction force properly. We begin by obtaining the differential forms of the conservation equations and property relations. Continuity equation The differential form of the continuity equation is obtained by differentiating the continuity relation rV ! constant and rearranging, Differential control volume dFfriction
x
P T
P " dP T " dT
V r
V " dV r " dr dx
r dV " V dr ! 0
→
dr dV !$ r V
(12–76)
. . . x-Momentum equation Noting that m1 ! m2 ! m ! rAV and A1 ! A2 ! A, applying the momentum equation → # → # → a F ! a bmV $ a bmV to the differential control volume out
in
in Fig. 12–62 gives # # PA c $ (P " dP)A $ dFfriction ! m (V " dV) $ mV
A1 ! A 2 ! A
FIGURE 12–62 Differential control volume for adiabatic flow in a constant-area duct with friction.
where we have again approximated the momentum flux correction factor b as 1. This equation simplifies to $dPA $ dFfriction ! rAV dV
or
dP "
dFfriction " rV dV ! 0 A
(12–77)
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661 CHAPTER 12
The friction force is related to the wall shear stress tw and the local friction factor fx by fx f x A dx 4A dx ! rV 2 dFfriction ! tw dA s ! tw p dx ! a rV 2b 8 Dh 2 Dh
(12–78)
where dx is the length of the flow section, p is the perimeter, and Dh ! 4A/p is the hydraulic diameter of the duct (note that Dh reduces to ordinary diameter D for a duct of circular cross section). Substituting, dP "
rV 2 f x dx " rV dV ! 0 2D h
(12–79)
Noting that V ! Ma1kRT and P ! rRT, we have rV 2 ! rkRTMa2 ! kPMa2 and rV ! kPMa2/V. Substituting into Eq. 12–79, fx dV 1 dP " !0 dx " 2D h V kMa2 P
(12–80)
Energy equation Noting that cp ! kR/(k $ 1) and V 2 ! Ma2kRT, the energy equation T0 ! constant or T " V 2/2cp ! constant can be expressed as T0 ! T a1 "
k$1 Ma2b ! constant 2
(12–81)
Differentiating and rearranging give
2(k $ 1)Ma2 dMa dT !$ T 2 " (k $ 1)Ma2 Ma
(12–82)
which is an expression for the differential change in temperature in terms of a differential change in Mach number. Mach number The Mach number relation for ideal gases can be expressed as V 2 ! Ma2kRT. Differentiating and rearranging give 2V dV ! 2MakRT dMa " kRMa2 dT →
(12–83)
V2 V2 2V dV ! 2 dMa " dT Ma T
Dividing each term by 2V 2 and rearranging, dV dMa 1 dT ! " V Ma 2 T
(12–84)
Combining Eq. 12–84 with Eq. 12–82 gives the velocity change in terms of the Mach number as (k $ 1)Ma2 dMa dV dMa ! $ V Ma 2 " (k $ 1)Ma2 Ma
dMa dV 2 (12–85) ! V 2 " (k $ 1)Ma2 Ma
or
Ideal gas The differential form of the ideal-gas equation is obtained by differentiating the equation P ! rRT, dP ! rR dT " RT dr
→
dP dT dr ! " r P T
(12–86)
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662 FLUID MECHANICS
Combining with the continuity equation (Eq. 12–76) gives dP dT dV ! $ P T V
(12–87)
Now combining with Eqs. 12–82 and 12–84 gives 2 " 2(k $ 1)Ma2 dMa dP !$ P 2 " (k $ 1)Ma2 Ma
(12–88)
which is an expression for differential changes in P with Ma. Substituting Eqs. 12–85 and 12–88 into 12–80 and simplifying give the differential equation for the variation of the Mach number with x as fx 4(1 $ Ma2) dx ! dMa 3 Dh kMa 32 " (k $ 1)Ma24
Sonic state as reference point
L
Ma* ! 1 P* T* V*
Ma P T V L* x
Hypothetical duct extension to sonic state
FIGURE 12–63 The length L* represents the distance between a given section where the Mach number is Ma and a real or imaginary section where Ma* ! 1.
(12–89)
Considering that all Fanno flows tend to Ma ! 1, it is again convenient to use the critical point (i.e., the sonic state) as the reference point and to express flow properties relative to the critical point properties, even if the actual flow never reaches the critical point. Integrating Eq. 12–89 from any state (Ma ! Ma and x ! x) to the critical state (Ma ! 1 and x ! xcr) gives (k " 1)Ma2 fL* 1 $ Ma2 k " 1 ! " ln 2 2k Dh kMa 2 " (k $ 1)Ma2
(12–90)
where f is the average friction factor between x and xcr, which is assumed to be constant, and L* ! xcr $ x is the channel length required for the Mach number to reach unity under the influence of wall friction. Therefore, L* represents the distance between a given section where the Mach number is Ma and a section (an imaginary section if the duct is not long enough to reach Ma ! 1) where sonic conditions occur (Fig. 12–63). Note that the value of fL*/Dh is fixed for a given Mach number, and thus values of fL*/Dh can be tabulated versus Ma for a specified k. Also, the value of duct length L* needed to reach sonic conditions (or the “sonic length”) is inversely proportional to the friction factor. Therefore, for a given Mach number, L* is large for ducts with smooth surfaces and small for ducts with rough surfaces. The actual duct length L between two sections where the Mach numbers are Ma1 and Ma2 can be determined from fL fL* fL* !a b $a b Dh Dh 1 Dh 2
(12–91)
The average friction factor f, in general, is different in different parts of the duct. If f is assumed to be constant for the entire duct (including the hypothetical extension part to the sonic state), then Eq. 12–91 simplifies to L ! L*1 $ L*2
( f ! constant)
(12–92)
Therefore, Eq. 12–90 can be used for short ducts that never reach Ma ! 1 as well as long ones with Ma ! 1 at the exit. The friction factor depends on the Reynolds number Re ! rVDh /m, which varies along the duct, and the roughness ratio e/Dh of the surface. The variation of Re is mild, however, since rV ! constant (from continuity), and any change in Re is due to the variation of viscosity with temperature.
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663 CHAPTER 12
Therefore, it is reasonable to evaluate f from the Moody chart or Colebrook equation discussed in Chap. 8 at the average Reynolds number and to treat it as a constant. This is the case for subsonic flow since the temperature changes involved are relatively small. The treatment of the friction factor for supersonic flow is beyond the scope of this text. The Colebrook equation is implicit in f, and thus it is more convenient to use the explicit Haaland relation expressed as 1 2f
6.9 e/D 1.11 "a b d Re 3.7
" $1.8 logc
(12–93)
The Reynolds numbers encountered in compressible flow are typically high, and at very high Reynolds numbers (fully rough turbulent flow) the friction factor is independent of the Reynolds number. For Re → *, the Colebrook equation reduces to 1/1f ! $2.0 log 3(e/D h)/3.74. Relations for other flow properties can be determined similarly by integrating the dP/P, dT/T, and dV/V relations from Eqs. 12–79, 12–82, and 12–85, respectively, from any state (no subscript and Mach number Ma) to the sonic state (with a superscript asterisk and Ma ! 1) with the following results (Fig. 12–64): (12–94)
T k"1 ! T* 2 " (k $ 1)Ma2
(12–95) 1/2
(12–96)
A similar relation can be obtained for the dimensionless stagnation pressure as follows: k/(k$1) 1/2 P0 P0 P P* k$1 1 k"1 k $ 1 $k/(k$1) 2 ! ! a1 " b b a1 " Ma a b 2 P*0 P P* P*0 2 Ma 2 " (k $ 1)Ma 2
which simplifies to
r0 P0 1 2 " (k $ 1)Ma2 (k"1)/[2(k$1)] ! *! a b * P 0 r 0 Ma k"1
(12–97)
Note that the stagnation temperature T0 is constant for Fanno flow, and thus T0/T *0 ! 1 everywhere along the duct. Eqs. 12–90 through 12–97 enable us to calculate the dimensionless pressure, temperature, density, velocity, stagnation pressure, and fL*/Dh for Fanno flow of an ideal gas with a specified k for any given Mach number. Representative results are given in tabular and graphical form in Table A–16 for k ! 1.4.
Choked Fanno Flow
!
It is clear from the previous discussions that friction causes subsonic Fanno flow in a constant-area duct to accelerate toward sonic velocity, and the Mach number becomes exactly unity at the exit for a certain duct length. This duct length is referred to as the maximum length, the sonic length, or the critical length, and is denoted by L*. You may be curious to know what happens if we extend the duct length beyond L*. In particular, does the flow accelerate to
r0 1 2 " (k $ 1)Ma2 (k"1)/[2(k$1)] a b ! r*0 Ma k"1
T k"1 ! T* 2 " (k $ 1)Ma2
1/2 P 1 k"1 b ! a P* Ma 2 " (k $ 1)Ma2
1/2 r* k"1 V ! ! Ma a b V* r 2 " (k $ 1)Ma2
fL*
1/2 P 1 k"1 ! a b P* Ma 2 " (k $ 1)Ma2
r* k"1 V ! Ma a ! b r V* 2 " (k $ 1)Ma2
P0 P0*
Dh
!
(k " 1)Ma2 1 $ Ma2 k " 1 " ln 2 2k kMa 2 " (k $ 1)Ma2
FIGURE 12–64 Summary of relations for Fanno flow.
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664 FLUID MECHANICS
Duct inlet
Duct exit Ma ! 1
Ma ( 1 Converging– diverging nozzle
Ma ' 1
Normal shock
FIGURE 12–65 Supersonic Fanno flow is always sonic at the duct exit. Extending the duct will only move the location of the normal shock further upstream.
supersonic velocities? The answer to this question is a definite no since at Ma ! 1 the flow is at the point of maximum entropy, and proceeding along the Fanno line to the supersonic region would require the entropy of the fluid to decrease—a violation of the second law of thermodynamics. (Note that the exit state must remain on the Fanno line to satisfy all conservation requirements.) Therefore, the flow is choked. This again is analogous to not being able to accelerate a gas to supersonic velocities in a converging nozzle by simply extending the converging flow section. If we extend the duct length beyond L* anyway, we simply move the critical state further downstream and reduce the flow rate. This causes the inlet state to change (e.g., inlet velocity decreases), and the flow shifts to a different Fanno line. Further increase in duct length further decreases the inlet velocity and thus the mass flow rate. Friction causes supersonic Fanno flow in a constant-area duct to decelerate and the Mach number to decrease toward unity. Therefore, the exit Mach number again becomes Ma ! 1 if the duct length is L*, as in subsonic flow. But unlike subsonic flow, increasing the duct length beyond L* cannot choke the flow since it is already choked. Instead, it causes a normal shock to occur at such a location that the continuing subsonic flow becomes sonic again exactly at the duct exit (Fig. 12–65). As the duct length increases, the location of the normal shock moves further upstream. Eventually, the shock occurs at the duct inlet. Further increase in duct length moves the shock to the diverging section of the converging–diverging nozzle that originally generates the supersonic flow, but the mass flow rate still remains unaffected since the mass flow rate is fixed by the sonic conditions at the throat of the nozzle, and it does not change unless the conditions at the throat change. EXAMPLE 12–16
Ma2 ! 1 P1 ! 150 kPa T1 ! 300 K
D ! 3 cm
Ma1 ! 0.4 L*1
FIGURE 12–66 Schematic for Example 12–16.
T* P* V*
Choked Fanno Flow in a Duct
Air enters a 3-cm-diameter smooth adiabatic duct at Ma1 ! 0.4, T1 ! 300 K, and P1 ! 150 kPa (Fig. 12–66). If the Mach number at the duct exit is 1, determine the duct length and temperature, pressure, and velocity at the duct exit. Also determine the percentage of stagnation pressure lost in the duct.
SOLUTION Air enters a constant-area adiabatic duct at a specified state and leaves at the sonic state. The duct length, exit temperature, pressure, velocity, and the percentage of stagnation pressure lost in the duct are to be determined. Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor is constant along the duct. Properties We take the properties of air to be k ! 1.4, cp ! 1.005 kJ/kg · K, R ! 0.287 kJ/kg · K, and n ! 1.58 # 10$5 m2/s. Analysis We first determine the inlet velocity and the inlet Reynolds number, c1 ! 2kRT1 !
B
(1.4)(0.287 kJ/kg % K)(300 K)a
1000 m2/s2 b ! 347 m/s 1 kJ/kg
V1 ! Ma1c1 ! 0.4(347 m/s) ! 139 m/s Re1 !
V1D (139 m/s)(0.03 m) ! ! 2.637 # 10 5 n 1.58 # 10 $5 m2/s
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665 CHAPTER 12
The friction factor is determined from the Colebrook equation,
1 2f
2.51 2.51 e/D 1 0 " " b → ! $2.0 loga b 3.7 Re 2f 3.7 2f 2.637 # 10 5 2f
! $2.0 log a
Its solution is
f ! 0.0148 The Fanno flow functions corresponding to the inlet Mach number of 0.4 are (Table A–16):
P01 fL*1 T1 P1 V1 ! 1.5901 ! 1.1628 ! 2.6958 ! 0.4313 ! 2.3085 * P0 T* P* V* D Noting that * denotes sonic conditions, which exist at the exit state, the duct length and the exit temperature, pressure, and velocity are determined to be
L*1 !
2.3085D 2.3085(0.03 m) ! ! 4.68 m f 0.0148
T* !
T1 300 K ! ! 258 K 1.1628 1.1628
P* !
P1 150 kPa ! ! 55.6 kPa 2.6958 2.6958
V* !
V1 139 m/s ! ! 322 m/s 0.4313 0.4313
Thus, for the given friction factor, the duct length must be 4.68 m for the Mach number to reach Ma ! 1 at the duct exit. The fraction of inlet stagnation pressure P01 lost in the duct due to friction is
P *0 P01 $ P*0 1 !1$ !1$ ! 0.371 or 37.1% P01 P01 1.5901 Discussion This problem can also be solved using appropriate relations instead of tabulated values for the Fanno functions. Also, we determined the friction factor at the inlet conditions and assumed it to remain constant along the duct. To check the validity of this assumption, we calculate the friction factor at the outlet conditions. It can be shown that the friction factor at the duct outlet is 0.0121—a drop of 18 percent, which is large. Therefore, we should repeat the calculation using the average value of the friction factor (0.0148 " 0.0121)/2 ! 0.0135. This would give the duct length to be L*1 ! 2.3085(0.03m)/0.0135 ! 5.13 m, and we take this to be the required duct length. L *2
L ! 27 m
EXAMPLE 12–17
Exit Conditions of Fanno Flow in a Duct
Air enters a 27-m-long 5-cm-diameter adiabatic duct at V1 ! 85 m/s, T1 ! 450 K, and P1 ! 220 kPa (Fig. 12–67). The average friction factor for the duct is estimated to be 0.023. Determine the Mach number at the duct exit and the mass flow rate of air.
SOLUTION Air enters a constant-area adiabatic duct of given length at a specified state. The exit Mach number and the mass flow rate are to be determined.
P1 ! 220 kPa T1 ! 450 K V1 ! 85 m/s
Exit Ma2
Ma* ! 1 T* P* V*
L*1 x
Hypothetical duct extension to sonic state
FIGURE 12–67 Schematic for Example 12–17.
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666 FLUID MECHANICS
Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor is constant along the duct. Properties We take the properties of air to be k ! 1.4, cp ! 1.005 kJ/kg · K, and R ! 0.287 kJ/kg · K. Analysis The first thing we need to know is whether the flow is choked at the exit or not. Therefore, we first determine the inlet Mach number and the corresponding value of the function fL*/Dh,
c1 ! 2kRT1 ! Ma1 !
B
1000 m2/s2 b ! 425 m/s 1 kJ/kg
(1.4)(0.287 kJ/kg % K)(450 K)a
V1 85 m/s ! ! 0.200 c1 425 m/s
Corresponding to this Mach number we read, from Table A–16, (fL*/Dh)1 ! 14.5333. Also, using the actual duct length L, we have
fL (0.023)(27 m) ! ! 12.42 ' 14.5333 Dh 0.05 m Therefore, flow is not choked and the exit Mach number is less than 1. The function fL*/Dh at the exit state is calculated from Eq. 12–91,
fL* fL* fL b !a b $ ! 14.5333 $ 12.42 ! 2.1133 Dh 2 Dh 1 Dh
a
The Mach number corresponding to this value of fL*/D is 0.42, obtained from Table A–16. Therefore, the Mach number at the duct exit is
Ma2 ! 0.42 The mass flow rate of air is determined from the inlet conditions to be
r1 !
P1 220 kPa 1 kJ ! b ! 1.703 kg/m3 a RT1 (0.287 kJ/kg % K)(450 K) 1 kPa % m3
# m air ! r 1A 1V1 ! (1.703 kg/m3) 3p(0.05 m)2/44 (85 m/s) ! 0.284 kg/s
Discussion Note that it takes a duct length of 27 m for the Mach number to increase from 0.20 to 0.42, but only 4.6 m to increase from 0.42 to 1. Therefore, the Mach number rises at a much higher rate as sonic conditions are approached. To gain some insight, let’s determine the lengths corresponding to fL*/Dh values at the inlet and the exit states. Noting that f is assumed to be constant for the entire duct, the maximum (or sonic) duct lengths at the inlet and exit states are
L max, 1 ! L*1 ! 14.5333 L max, 2 ! L*2 ! 2.1133
Dh 0.05 m ! 14.5333 ! 31.6 m f 0.023
Dh 0.05 m ! 2.1133 ! 4.59 m f 0.023
(or, Lmax, 2 ! Lmax, 1 $ L ! 31.6 $ 27 ! 4.6 m). Therefore, the flow would reach sonic conditions if a 4.6-m-long section were added to the existing duct.
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APPLICATION SPOTLIGHT
■
Shock-Wave/Boundary-Layer Interactions
Guest Author: Gary S. Settles, The Pennsylvania State University Shock waves and boundary layers are among nature’s most incompatible phenomena. Boundary layers, as described in Chap. 10, are susceptible to separation from aerodynamic surfaces wherever strong adverse pressure gradients occur. Shock waves, on the other hand, produce very strong adverse pressure gradients, since a finite rise in static pressure occurs across a shock wave over a negligibly short streamwise distance. Thus, when a boundary layer encounters a shock wave, a complicated flow pattern develops and the boundary layer often separates from the surface to which it was attached. There are important cases in high-speed flight and wind tunnel testing where such a clash is unavoidable. For example, commercial jet transport aircraft cruise in the bottom edge of the transonic flow regime, where the airflow over their wings actually goes supersonic and then returns to subsonic flow through a normal shock wave (Fig. 12–68). If such an aircraft flies significantly faster than its design cruise Mach number, serious aerodynamic disturbances arise due to shock-wave/boundary-layer interactions causing flow separation on the wings. This phenomenon thus limits the speed of passenger aircraft around the world. Some military aircraft are designed to avoid this limit and fly supersonically, but shock-wave/boundary-layer interactions are still limiting factors in their engine air inlets. The interaction of a shock wave and a boundary layer is a type of viscous–inviscid interaction in which the viscous flow in the boundary layer encounters the essentially inviscid shock wave generated in the free stream. The boundary layer is slowed and thickened by the shock and may separate. The shock, on the other hand, bifurcates when flow separation occurs (Fig. 12–69). Mutual changes in both the shock and the boundary layer continue until an equilibrium condition is reached. Depending upon boundary conditions, the interaction can vary in either two or three dimensions and may be steady or unsteady. Such a strongly interacting flow is difficult to analyze, and no simple solutions exist. Moreover, in most of the problems of practical interest, the boundary layer in question is turbulent. Modern computational methods are able to predict many features of these flows by supercomputer solutions of the Reynolds-averaged Navier–Stokes equations. Wind tunnel experiments play a key role in guiding and validating such computations. Overall, the shock-wave/boundary-layer interaction has become one of the pacing problems of modern fluid dynamics research. References Knight, D. D., et al., “Advances in CFD Prediction of Shock Wave Turbulent Boundary Layer Interactions,” Progress in Aerospace Sciences 39(2-3), pp. 121–184, 2003. Alvi, F. S., and Settles, G. S., “Physical Model of the Swept Shock Wave/Boundary-Layer Interaction Flowfield,” AIAA Journal 30, pp. 2252–2258, Sept. 1992.
FIGURE 12–68 Normal shock wave above the wing of an L-1011 commercial jet aircraft in transonic flight, made visible by background distortion of low clouds over the Pacific Ocean. U.S. Govt. photo by Carla Thomas, NASA Dryden Research Center.
FIGURE 12–69 Shadowgram of the swept interaction generated by a fin mounted on a flat plate at Mach 3.5. The oblique shock wave generated by the fin (at top of image) bifurcates into a “l-foot” beneath which the boundary layer separates and rolls up. The airflow through the l-foot above the separation zone forms a supersonic “jet” that curves downward and impinges upon the wall. This three-dimensional interaction required a special optical technique known as conical shadowgraphy to visualize the flow. Photo by F. S. Alvi and G. S. Settles.
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SUMMARY In this chapter the effects of compressibility on gas flow are examined. When dealing with compressible flow, it is convenient to combine the enthalpy and the kinetic energy of the fluid into a single term called stagnation (or total) enthalpy h0, defined as h0 ! h "
V2 2
and
The properties of a fluid at the stagnation state are called stagnation properties and are indicated by the subscript zero. The stagnation temperature of an ideal gas with constant specific heats is T0 ! T "
V2 2cp
r0 T0 1/(k$1) !a b r T
and
The velocity at which an infinitesimally small pressure wave travels through a medium is the speed of sound. For an ideal gas it is expressed as c!
&P a b ! 2kRT B &r s
The Mach number is the ratio of the actual velocity of the fluid to the speed of sound at the same state: Ma !
1/(k$1) r0 k$1 ! c1 " a b Ma2d r 2
T* 2 ! T0 k " 1 1/(k$1) r* 2 !a b r0 k"1
P* 2 k/(k$1) !a b P0 k"1
The pressure outside the exit plane of a nozzle is called the back pressure. For all back pressures lower than P*, the pressure at the exit plane of the converging nozzle is equal to P*, the Mach number at the exit plane is unity, and the mass flow rate is the maximum (or choked) flow rate. In some range of back pressure, the fluid that achieved a sonic velocity at the throat of a converging–diverging nozzle and is accelerating to supersonic velocities in the diverging section experiences a normal shock, which causes a sudden rise in pressure and temperature and a sudden drop in velocity to subsonic levels. Flow through the shock is highly irreversible, and thus it cannot be approximated as isentropic. The properties of an ideal gas with constant specific heats before (subscript 1) and after (subscript 2) a shock are related by T01 ! T02
V c
The flow is called sonic when Ma ! 1, subsonic when Ma ' 1, supersonic when Ma ( 1, hypersonic when Ma (( 1, and transonic when Ma ! 1. Nozzles whose flow area decreases in the flow direction are called converging nozzles. Nozzles whose flow area first decreases and then increases are called converging–diverging nozzles. The location of the smallest flow area of a nozzle is called the throat. The highest velocity to which a fluid can be accelerated in a converging nozzle is the sonic velocity. Accelerating a fluid to supersonic velocities is possible only in converging–diverging nozzles. In all supersonic converging–diverging nozzles, the flow velocity at the throat is the velocity of sound. The ratios of the stagnation to static properties for ideal gases with constant specific heats can be expressed in terms of the Mach number as
k/(k$1) P0 k$1 ! c1 " a b Ma2d P 2
When Ma ! 1, the resulting static-to-stagnation property ratios for the temperature, pressure, and density are called critical ratios and are denoted by the superscript asterisk:
and
which represents the temperature an ideal gas would attain if it is brought to rest adiabatically. The stagnation properties of an ideal gas are related to the static properties of the fluid by T0 k/(k$1) P0 !a b P T
T0 k$1 !1" a b Ma2 T 2
Ma2 !
(k $ 1)Ma21 " 2 B 2kMa21 $ k " 1
T2 2 " Ma21(k $ 1) ! T1 2 " Ma22(k $ 1) and
P2 1 " kMa21 2kMa21 $ k " 1 ! ! P1 1 " kMa22 k"1
These equations also hold across an oblique shock, provided that the component of the Mach number normal to the oblique shock is used in place of the Mach number. Steady one-dimensional flow of an ideal gas with constant specific heats through a constant-area duct with heat transfer and negligible friction is referred to as Rayleigh flow. The property relations and curves for Rayleigh flow are given in Table A–15. Heat transfer during Rayleigh flow can be determined from q ! cp(T02 $ T01) ! cp(T2 $ T1) "
V 22 $ V 21 2
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669 CHAPTER 12
Steady, frictional, and adiabatic flow of an ideal gas with constant specific heats through a constant-area duct is referred to as Fanno flow. The channel length required for the Mach number to reach unity under the influence of wall friction is denoted by L* and is expressed as fL* 1 $ Ma2 k " 1 (k " 1)Ma2 ! " ln Dh 2k kMa2 2 " (k $ 1)Ma2 where f is the average friction factor. The duct length between two sections where the Mach numbers are Ma1 and Ma2 is determined from
fL fL* fL* !a b $a b Dh Dh 1 Dh 2
During Fanno flow, the stagnation temperature T0 remains constant. Other property relations and curves for Fanno flow are given in Table A–16. This chapter provides an overview of compressible flow and is intended to motivate the interested student to undertake a more in-depth study of this exciting subject. Some compressible flows are analyzed in Chap. 15 using computational fluid dynamics.
REFERENCES AND SUGGESTED READING 1. J. D. Anderson. Modern Compressible Flow with Historical Perspective, 3rd ed. New York: McGraw-Hill, 2003. 2. Y. A. Çengel and M. A. Boles. Thermodynamics: An Engineering Approach, 4th ed. New York: McGraw-Hill, 2002. 3. H. Cohen, G. F. C. Rogers, and H. I. H. Saravanamuttoo. Gas Turbine Theory, 3rd ed. New York: Wiley, 1987. 4. W. J. Devenport. Compressible Aerodynamic Calculator, http://www.aoe.vt.edu/~devenpor/aoe3114/calc.html. 5. R. W. Fox and A. T. McDonald. Introduction to Fluid Mechanics, 5th ed. New York: Wiley, 1999. 6. H. Liepmann and A. Roshko. Elements of Gas Dynamics, Dover Publications, Mineola, NY, 2001.
7. C. E. Mackey, responsible NACA officer and curator. Equations, Tables, and Charts for Compressible Flow. NACA Report 1135, http://naca.larc.nasa.gov/reports/ 1953/naca-report-1135/. 8. A. H. Shapiro. The Dynamics and Thermodynamics of Compressible Fluid Flow, vol. 1. New York: Ronald Press Company, 1953. 9. P. A. Thompson. Compressible-Fluid Dynamics, New York: McGraw-Hill, 1972. 10. United Technologies Corporation. The Aircraft Gas Turbine and Its Operation, 1982. 11. Van Dyke, 1982. 12. F. M. White. Fluid Mechanics, 5th ed. New York: McGraw-Hill, 2003.
PROBLEMS* Stagnation Properties 12–1C A high-speed aircraft is cruising in still air. How does the temperature of air at the nose of the aircraft differ from the temperature of air at some distance from the aircraft? 12–2C How and why is the stagnation enthalpy h0 defined? How does it differ from ordinary (static) enthalpy? 12–3C
What is dynamic temperature?
12–4C In air-conditioning applications, the temperature of air is measured by inserting a probe into the flow stream. * Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
Thus, the probe actually measures the stagnation temperature. Does this cause any significant error? 12–5 Determine the stagnation temperature and stagnation pressure of air that is flowing at 44 kPa, 245.9 K, and 470 m/s. Answers: 355.8 K, 160.3 kPa
12–6 Air at 300 K is flowing in a duct at a velocity of (a) 1, (b) 10, (c) 100, and (d) 1000 m/s. Determine the temperature that a stationary probe inserted into the duct will read for each case. 12–7 Calculate the stagnation temperature and pressure for the following substances flowing through a duct: (a) helium at 0.25 MPa, 50°C, and 240 m/s; (b) nitrogen at 0.15 MPa, 50°C, and 300 m/s; and (c) steam at 0.1 MPa, 350°C, and 480 m/s. 12–8 Air enters a compressor with a stagnation pressure of 100 kPa and a stagnation temperature of 27°C, and it is compressed to a stagnation pressure of 900 kPa. Assuming the
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compression process to be isentropic, determine the power input to the compressor for a mass flow rate of 0.02 kg/s.
12–22 Assuming ideal gas behavior, determine the speed of sound in refrigerant-134a at 0.1 MPa and 60°C.
Answer: 5.27 kW
12–23 The Airbus A-340 passenger plane has a maximum takeoff weight of about 260,000 kg, a length of 64 m, a wing span of 60 m, a maximum cruising speed of 945 km/h, a seating capacity of 271 passengers, maximum cruising altitude of 14,000 m, and a maximum range of 12,000 km. The air temperature at the crusing altitude is about $60°C. Determine the Mach number of this plane for the stated limiting conditions.
12–9E Steam flows through a device with a stagnation pressure of 120 psia, a stagnation temperature of 700°F, and a velocity of 900 ft/s. Assuming ideal-gas behavior, determine the static pressure and temperature of the steam at this state. 12–10 Products of combustion enter a gas turbine with a stagnation pressure of 1.0 MPa and a stagnation temperature of 750°C, and they expand to a stagnation pressure of 100 kPa. Taking k ! 1.33 and R ! 0.287 kJ/kg · K for the products of combustion, and assuming the expansion process to be isentropic, determine the power output of the turbine per unit mass flow.
12–24E Steam flows through a device with a pressure of 120 psia, a temperature of 700°F, and a velocity of 900 ft/s. Determine the Mach number of the steam at this state by assuming ideal-gas behavior with k ! 1.3. Answer: 0.441
12–11 Air flows through a device such that the stagnation pressure is 0.6 MPa, the stagnation temperature is 400°C, and the velocity is 570 m/s. Determine the static pressure and temperature of the air at this state. Answers: 518.6 K, 0.23 MPa
12–25E
Speed of Sound and Mach Number
12–26 The isentropic process for an ideal gas is expressed as Pv k ! constant. Using this process equation and the definition of the speed of sound (Eq. 12–9), obtain the expression for the speed of sound for an ideal gas (Eq. 12–11).
12–12C What is sound? How is it generated? How does it travel? Can sound waves travel in a vacuum? 12–13C Is it realistic to assume that the propagation of sound waves is an isentropic process? Explain. 12–14C Is the sonic velocity in a specified medium a fixed quantity, or does it change as the properties of the medium change? Explain.
Reconsider Prob. 12–24E. Using EES (or other) software, compare the Mach number of steam flow over the temperature range 350 to 700°F. Plot the Mach number as a function of temperature.
12–27 Air expands isentropically from 1.5 MPa and 60°C to 0.4 MPa. Calculate the ratio of the initial to final speed of sound. Answer: 1.21 12–28
Repeat Prob. 12–27 for helium gas.
12–15C In which medium does a sound wave travel faster: in cool air or in warm air?
12–29E Air expands isentropically from 170 psia and 200°F to 60 psia. Calculate the ratio of the initial to final speed of sound. Answer: 1.16
12–16C In which medium will sound travel fastest for a given temperature: air, helium, or argon?
One-Dimensional Isentropic Flow
12–17C In which medium does a sound wave travel faster: in air at 20°C and 1 atm or in air at 20°C and 5 atm? 12–18C Does the Mach number of a gas flowing at a constant velocity remain constant? Explain. 12–19 Determine the speed of sound in air at (a) 300 K and (b) 1000 K. Also determine the Mach number of an aircraft moving in air at a velocity of 240 m/s for both cases. 12–20 Carbon dioxide enters an adiabatic nozzle at 1200 K with a velocity of 50 m/s and leaves at 400 K. Assuming constant specific heats at room temperature, determine the Mach number (a) at the inlet and (b) at the exit of the nozzle. Assess the accuracy of the constant specific heat assumption. Answers: (a) 0.0925, (b) 3.73
12–21 Nitrogen enters a steady-flow heat exchanger at 150 kPa, 10°C, and 100 m/s, and it receives heat in the amount of 120 kJ/kg as it flows through it. Nitrogen leaves the heat exchanger at 100 kPa with a velocity of 200 m/s. Determine the Mach number of the nitrogen at the inlet and the exit of the heat exchanger.
12–30C Consider a converging nozzle with sonic speed at the exit plane. Now the nozzle exit area is reduced while the nozzle inlet conditions are maintained constant. What will happen to (a) the exit velocity and (b) the mass flow rate through the nozzle? 12–31C A gas initially at a supersonic velocity enters an adiabatic converging duct. Discuss how this affects (a) the velocity, (b) the temperature, (c) the pressure, and (d) the density of the fluid. 12–32C A gas initially at a supersonic velocity enters an adiabatic diverging duct. Discuss how this affects (a) the velocity, (b) the temperature, (c) the pressure, and (d) the density of the fluid. 12–33C A gas initially at a subsonic velocity enters an adiabatic converging duct. Discuss how this affects (a) the velocity, (b) the temperature, (c) the pressure, and (d) the density of the fluid. 12–34C A gas initially at a subsonic velocity enters an adiabatic diverging duct. Discuss how this affects (a) the veloc-
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ity, (b) the temperature, (c) the pressure, and (d) the density of the fluid.
velocity, (b) the exit pressure, and (c) the mass flow rate through the nozzle?
12–35C A gas at a specified stagnation temperature and pressure is accelerated to Ma ! 2 in a converging–diverging nozzle and to Ma ! 3 in another nozzle. What can you say about the pressures at the throats of these two nozzles?
12–47C Consider a converging nozzle and a converging– diverging nozzle having the same throat areas. For the same inlet conditions, how would you compare the mass flow rates through these two nozzles?
12–36C Is it possible to accelerate a gas to a supersonic velocity in a converging nozzle?
12–48C Consider gas flow through a converging nozzle with specified inlet conditions. We know that the highest velocity the fluid can have at the nozzle exit is the sonic velocity, at which point the mass flow rate through the nozzle is a maximum. If it were possible to achieve hypersonic velocities at the nozzle exit, how would it affect the mass flow rate through the nozzle?
12–37 Air enters a converging–diverging nozzle at a pressure of 1.2 MPa with negligible velocity. What is the lowest pressure that can be obtained at the throat of the nozzle? Answer: 634 kPa
12–38 Helium enters a converging–diverging nozzle at 0.7 MPa, 800 K, and 100 m/s. What are the lowest temperature and pressure that can be obtained at the throat of the nozzle? 12–39 Calculate the critical temperature, pressure, and density of (a) air at 200 kPa, 100°C, and 250 m/s, and (b) helium at 200 kPa, 40°C, and 300 m/s.
12–49C How does the parameter Ma* differ from the Mach number Ma? 12–50C What would happen if we attempted to decelerate a supersonic fluid with a diverging diffuser? 12–51C What would happen if we tried to further accelerate a supersonic fluid with a diverging diffuser?
12–40 Quiescent carbon dioxide at 800 kPa and 400 K is accelerated isentropically to a Mach number of 0.6. Determine the temperature and pressure of the carbon dioxide after acceleration. Answers: 380 K, 636 kPa
12–52C Consider the isentropic flow of a fluid through a converging–diverging nozzle with a subsonic velocity at the throat. How does the diverging section affect (a) the velocity, (b) the pressure, and (c) the mass flow rate of the fluid?
12–41 Air at 200 kPa, 100°C, and Mach number Ma ! 0.8 flows through a duct. Calculate the velocity and the stagnation pressure, temperature, and density of the air.
12–53C Is it possible to accelerate a fluid to supersonic velocities with a velocity other than the sonic velocity at the throat? Explain
12–42
12–54 Explain why the maximum flow rate per unit area for a given ideal gas depends only on P0 / 1T0. For an ideal gas with k ! 1.4 and R ! 0.287 kJ/kg · K, find the constant a # such that m/A* ! aP0 /1T 0.
Reconsider Prob. 12–41. Using EES (or other) software, study the effect of Mach numbers in the range 0.1 to 2 on the velocity, stagnation pressure, temperature, and density of air. Plot each parameter as a function of the Mach number. 12–43E Air at 30 psia, 212°F, and Mach number Ma ! 0.8 flows through a duct. Calculate the velocity and the stagnation pressure, temperature, and density of air. Answers:
1016 ft/s, 758 R, 45.7 psia, 0.163 lbm/ft3
12–44 An aircraft is designed to cruise at Mach number Ma ! 1.4 at 8000 m where the atmospheric temperature is 236.15 K. Determine the stagnation temperature on the leading edge of the wing.
Isentropic Flow through Nozzles 12–45C Consider subsonic flow in a converging nozzle with fixed inlet conditions. What is the effect of dropping the back pressure to the critical pressure on (a) the exit velocity, (b) the exit pressure, and (c) the mass flow rate through the nozzle? 12–46C Consider subsonic flow in a converging nozzle with specified conditions at the nozzle inlet and critical pressure at the nozzle exit. What is the effect of dropping the back pressure well below the critical pressure on (a) the exit
12–55 For an ideal gas obtain an expression for the ratio of the speed of sound where Ma ! 1 to the speed of sound based on the stagnation temperature, c*/c0. 12–56 An ideal gas flows through a passage that first converges and then diverges during an adiabatic, reversible, steady-flow process. For subsonic flow at the inlet, sketch the variation of pressure, velocity, and Mach number along the length of the nozzle when the Mach number at the minimum flow area is equal to unity. 12–57
Repeat Prob. 12–56 for supersonic flow at the inlet.
12–58 Air enters a nozzle at 0.2 MPa, 350 K, and a velocity of 150 m/s. Assuming isentropic flow, determine the pressure and temperature of air at a location where the air velocity equals the speed of sound. What is the ratio of the area at this location to the entrance area? Answers: 301.0 K, 0.118 MPa, 0.629
12–59 Repeat Prob. 12–58 assuming the entrance velocity is negligible. 12–60E Air enters a nozzle at 30 psia, 630 R, and a velocity of 450 ft/s. Assuming isentropic flow, determine the pressure
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and temperature of air at a location where the air velocity equals the speed of sound. What is the ratio of the area at this location to the entrance area? Answers: 539 K, 17.4 psia, 0.574
12–61 Air enters a converging–diverging nozzle at 0.8 MPa with a negligible velocity. Assuming the flow to be isentropic, determine the back pressure that will result in an exit Mach number of 1.8. Answer: 0.139 MPa 12–62 Nitrogen enters a converging–diverging nozzle at 700 kPa and 400 K with a negligible velocity. Determine the critical velocity, pressure, temperature, and density in the nozzle. 12–63 An ideal gas with k ! 1.4 is flowing through a nozzle such that the Mach number is 2.4 where the flow area is 25 cm2. Assuming the flow to be isentropic, determine the flow area at the location where the Mach number is 1.2. 12–64 Repeat Prob. 12–63 for an ideal gas with k ! 1.33. 12–65 Air at 900 kPa and 400 K enters a converging nozzle with a negligible velocity. The throat area of the nozzle is 10 cm2. Assuming isentropic flow, calculate and plot the exit pressure, the exit velocity, and the mass flow rate versus the back pressure Pb for 0.9 + Pb + 0.1 MPa. 12–66
Reconsider Prob. 12–65. Using EES (or other) software, solve the problem for the inlet conditions of 1 MPa and 1000 K. 12–67E Air enters a converging–diverging nozzle of a supersonic wind tunnel at 150 psia and 100°F with a low velocity. The flow area of the test section is equal to the exit area of the nozzle, which is 5 ft2. Calculate the pressure, temperature, velocity, and mass flow rate in the test section for a Mach number Ma ! 2. Explain why the air must be very dry for this application. Answers: 19.1 psia, 311 R, 1729 ft/s, 1435 lbm/s
Shock Waves and Expansion Waves 12–68C Can a shock wave develop in the converging section of a converging–diverging nozzle? Explain. 12–69C What do the states on the Fanno line and the Rayleigh line represent? What do the intersection points of these two curves represent? 12–70C Can the Mach number of a fluid be greater than 1 after a normal shock wave? Explain. 12–71C How does the normal shock affect (a) the fluid velocity, (b) the static temperature, (c) the stagnation temperature, (d) the static pressure, and (e) the stagnation pressure? 12–72C How do oblique shocks occur? How do oblique shocks differ from normal shocks? 12–73C For an oblique shock to occur, does the upstream flow have to be supersonic? Does the flow downstream of an oblique shock have to be subsonic?
12–74C It is claimed that an oblique shock can be analyzed like a normal shock provided that the normal component of velocity (normal to the shock surface) is used in the analysis. Do you agree with this claim? 12–75C Consider supersonic airflow approaching the nose of a two-dimensional wedge and experiencing an oblique shock. Under what conditions does an oblique shock detach from the nose of the wedge and form a bow wave? What is the numerical value of the shock angle of the detached shock at the nose? 12–76C Consider supersonic flow impinging on the rounded nose of an aircraft. Will the oblique shock that forms in front of the nose be an attached or detached shock? Explain. 12–77C Are the isentropic relations of ideal gases applicable for flows across (a) normal shock waves, (b) oblique shock waves, and (c) Prandtl–Meyer expansion waves? 12–78 For an ideal gas flowing through a normal shock, develop a relation for V2/V1 in terms of k, Ma1, and Ma2. 12–79 Air enters a converging–diverging nozzle of a supersonic wind tunnel at 1 MPa and 300 K with a low velocity. If a normal shock wave occurs at the exit plane of the nozzle at Ma ! 2, determine the pressure, temperature, Mach number, velocity, and stagnation pressure after the shock wave. Answers: 0.575 MPa, 281 K, 0.577, 194 m/s, 0.721 MPa
12–80 Air enters a converging–diverging nozzle with low velocity at 2.0 MPa and 100°C. If the exit area of the nozzle is 3.5 times the throat area, what must the back pressure be to produce a normal shock at the exit plane of the nozzle? Answer: 0.661 MPa
12–81 What must the back pressure be in Prob. 12–80 for a normal shock to occur at a location where the cross-sectional area is twice the throat area? 12–82 Air flowing steadily in a nozzle experiences a normal shock at a Mach number of Ma ! 2.5. If the pressure and temperature of air are 61.64 kPa and 262.15 K, respectively, upstream of the shock, calculate the pressure, temperature, velocity, Mach number, and stagnation pressure downstream of the shock. Compare these results to those for helium undergoing a normal shock under the same conditions. 12–83 Calculate the entropy change of air across the normal shock wave in Prob. 12–82. 12–84E
Air flowing steadily in a nozzle experiences a normal shock at a Mach number of Ma ! 2.5. If the pressure and temperature of air are 10.0 psia and 440.5 R, respectively, upstream of the shock, calculate the pressure, temperature, velocity, Mach number, and stagnation pressure downstream of the shock. Compare these results to those for helium undergoing a normal shock under the same conditions.
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12–85E
Reconsider Prob. 12–84E. Using EES (or other) software, study the effects of both air and helium flowing steadily in a nozzle when there is a normal shock at a Mach number in the range 2 ' Ma1 ' 3.5. In addition to the required information, calculate the entropy change of the air and helium across the normal shock. Tabulate the results in a parametric table.
the flow direction. As a result, a weak oblique shock forms. Determine the wave angle, Mach number, pressure, and temperature after the shock.
12–86 Air enters a normal shock at 22.6 kPa, 217 K, and 680 m/s. Calculate the stagnation pressure and Mach number upstream of the shock, as well as pressure, temperature, velocity, Mach number, and stagnation pressure downstream of the shock.
12–95E Air flowing at P1 ! 6 psia, T1 ! 480 R, and Ma1 ! 2.0 is forced to undergo a compression turn of 15°. Determine the Mach number, pressure, and temperature of air after the compression.
12–94 Air flowing at P1 ! 40 kPa, T1 ! 280 K, and Ma1 ! 3.6 is forced to undergo an expansion turn of 15°. Determine the Mach number, pressure, and temperature of air after the expansion. Answers: 4.81, 8.31 kPa, 179 K
12–87 Calculate the entropy change of air across the normal shock wave in Problem 12–86. Answer: 0.155 kJ/kg · K
Duct Flow with Heat Transfer and Negligible Friction (Rayleigh Flow)
12–88
12–96C What is the characteristic aspect of Rayleigh flow? What are the main assumptions associated with Rayleigh flow?
Calculate and plot the entropy change of air across the normal shock for upstream Mach numbers between 0.5 and 1.5 in increments of 0.1. Explain why normal shock waves can occur only for upstream Mach numbers greater than Ma ! 1.
12–89 Consider supersonic airflow approaching the nose of a two-dimensional wedge at a Mach number of 5. Using Fig. 12–41, determine the minimum shock angle and the maximum deflection angle a straight oblique shock can have. 12–90 Air flowing at 60 kPa, 240 K, and a Mach number of 3.4 impinges on a two-dimensional wedge of half-angle 12°. Determine the two possible oblique shock angles, bweak and bstrong, that could be formed by this wedge. For each case, calculate the pressure, temperature, and Mach number downstream of the oblique shock. 12–91 Consider the supersonic flow of air at upstream conditions of 70 kPa and 260 K and a Mach number of 2.4 over a two-dimensional wedge of half-angle 10°. If the axis of the wedge is tilted 25° with respect to the upstream air flow, determine the downstream Mach number, pressure, and temperature above the wedge. Answers: 3.105, 23.8 kPa, 191 K
Ma 2 Ma1 ! 2.4 25°
10°
12–97C On a T-s diagram of Rayleigh flow, what do the points on the Rayleigh line represent? 12–98C What is the effect of heat gain and heat loss on the entropy of the fluid during Rayleigh flow? 12–99C Consider subsonic Rayleigh flow of air with a Mach number of 0.92. Heat is now transferred to the fluid and the Mach number increases to 0.95. Will the temperature T of the fluid increase, decrease, or remain constant during this process? How about the stagnation temperature T0? 12–100C What is the effect of heating the fluid on the flow velocity in subsonic Rayleigh flow? Answer the same questions for supersonic Rayleigh flow. 12–101C Consider subsonic Rayleigh flow that is accelerated to sonic velocity (Ma ! 1) at the duct exit by heating. If the fluid continues to be heated, will the flow at duct exit be supersonic, subsonic, or remain sonic? 12–102 Consider a 12-cm-diameter tubular combustion chamber. Air enters the tube at 500 K, 400 kPa, and 70 m/s. Fuel with a heating value of 39,000 kJ/kg is burned by spraying it into the air. If the exit Mach number is 0.8, determine the rate at which the fuel is burned and the exit temperature. Assume complete combustion and disregard the increase in the mass flow rate due to the fuel mass. Fuel P1 ! 400 kPa T1 ! 500 K
FIGURE P12–91 12–92 Reconsider Prob. 12–91. Determine the downstream Mach number, pressure, and temperature below the wedge for a strong oblique shock for an upstream Mach number of 5. 12–93E Air at 8 psia, 20°F, and a Mach number of 2.0 is forced to turn upward by a ramp that makes an 8° angle off
V1 ! 70 m/s
Ma2 ! 0.8 Combustor tube
FIGURE P12–102 12–103 Air enters a rectangular duct at T1 ! 300 K, P1 ! 420 kPa, and Ma1 ! 2. Heat is transferred to the air in the
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amount of 55 kJ/kg as it flows through the duct. Disregarding frictional losses, determine the temperature and Mach number at the duct exit. Answers: 386 K, 1.64 55 kJ/kg
Ma1 ! 1.8, P01 ! 210 kPa, and T01 ! 600 K, and it is decelerated by heating. Determine the highest temperature that air can be heated by heat addition while the mass flow rate remains constant.
Adiabatic Duct Flow with Friction (Fanno Flow) P1 ! 420 kPa T1 ! 300 K
Air
12–113C What is the characteristic aspect of Fanno flow? What are the main assumptions associated with Fanno flow?
Ma1 ! 2
12–114C On a T-s diagram of Fanno flow, what do the points on the Fanno line represent?
FIGURE P12–103
12–115C What is the effect of friction on the entropy of the fluid during Fanno flow?
12–104 Repeat Prob. 12–103 assuming air is cooled in the amount of 55 kJ/kg. 12–105 Air is heated as it flows subsonically through a duct. When the amount of heat transfer reaches 60 kJ/kg, the flow is observed to be choked, and the velocity and the static pressure are measured to be 620 m/s and 270 kPa. Disregarding frictional losses, determine the velocity, static temperature, and static pressure at the duct inlet. 12–106E Air flows with negligible friction through a 4-indiameter duct at a rate of 5 lbm/s. The temperature and pressure at the inlet are T1 ! 800 R and P1 ! 30 psia, and the Mach number at the exit is Ma2 ! 1. Determine the rate of heat transfer and the pressure drop for this section of the duct. 12–107
Air enters a frictionless duct with V1 ! 70 m/s, T1 ! 600 K, and P1 ! 350 kPa. Letting the exit temperature T2 vary from 600 to 5000 K, evaluate the entropy change at intervals of 200 K, and plot the Rayleigh line on a T-s diagram. 12–108E Air is heated as it flows through a 4 in # 4 in square duct with negligible friction. At the inlet, air is at T1 ! 700 R, P1 ! 80 psia, and V1 ! 260 ft/s. Determine the rate at which heat must be transferred to the air to choke the flow at the duct exit, and the entropy change of air during this process. 12–109 Compressed air from the compressor of a gas turbine enters the combustion chamber at T1 ! 550 K, P1 ! 600 kPa, and Ma1 ! 0.2 at a rate of 0.3 kg/s. Via combustion, heat is transferred to the air at a rate of 200 kJ/s as it flows through the duct with negligible friction. Determine the Mach number at the duct exit, and the drop in stagnation pressure P01 $ P02 during this process. Answers: 0.319, 21.8 kPa
12–116C Consider subsonic Fanno flow of air with an inlet Mach number of 0.70. If the Mach number increases to 0.90 at the duct exit as a result of friction, will the (a) stagnation temperature T0, (b) stagnation pressure P0, and (c) entropy s of the fluid increase, decrease, or remain constant during this process? 12–117C Consider supersonic Fanno flow of air with an inlet Mach number of 1.8. If the Mach number decreases to 1.2 at the duct exit as a result of friction, will the (a) stagnation temperature T0, (b) stagnation pressure P0, and (c) entropy s of the fluid increase, decrease, or remain constant during this process? 12–118C What is the effect of friction on flow velocity in subsonic Fanno flow? Answer the same question for supersonic Fanno flow. 12–119C Consider subsonic Fanno flow accelerated to sonic velocity (Ma ! 1) at the duct exit as a result of frictional effects. If the duct length is increased further, will the flow at the duct exit be supersonic, subsonic, or remain sonic? Will the mass flow rate of the fluid increase, decrease, or remain constant as a result of increasing the duct length? 12–120C Consider supersonic Fanno flow that is decelerated to sonic velocity (Ma ! 1) at the duct exit as a result of frictional effects. If the duct length is increased further, will the flow at the duct exit be supersonic, subsonic, or remain sonic? Will the mass flow rate of the fluid increase, decrease, or remain constant as a result of increasing the duct length? 12–121 Air enters a 5-cm-diameter adiabatic duct at Ma1 ! 0.2, T1 ! 400 K, and P1 ! 200 kPa. The average friction factor for the duct is estimated to be 0.016. If the Mach num-
12–110 Repeat Prob. 12–109 for a heat transfer rate of 300 kJ/s. 12–111 Argon gas enters a constant cross-sectional area duct at Ma1 ! 0.2, P1 ! 320 kPa, and T1 ! 400 K at a rate of 0.8 kg/s. Disregarding frictional losses, determine the highest rate of heat transfer to the argon without reducing the mass flow rate. 12–112 Consider supersonic flow of air through a 6-cmdiameter duct with negligible friction. Air enters the duct at
P1 ! 200 kPa T1 ! 400 K
Ma2 ! 0.8
Ma1 ! 0.2 L
FIGURE P12–121
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ber at the duct exit is 0.8, determine the duct length, temperature, pressure, and velocity at the duct exit. 12–122 Air enters a 15-m-long, 4-cm-diameter adiabatic duct at V1 ! 70 m/s, T1 ! 500 K, and P1 ! 300 kPa. The average friction factor for the duct is estimated to be 0.023. Determine the Mach number at the duct exit, the exit velocity, and the mass flow rate of air. 12–123 Air in a room at T0 ! 290 K and P0 ! 95 kPa is drawn steadily by a vacuum pump through a 1-cm-diameter, 50-cm-long adiabatic tube equipped with a converging nozzle at the inlet. The flow in the nozzle section can be assumed to be isentropic, and the average friction factor for the duct can be taken to be 0.018. Determine the maximum mass flow rate of air that can be sucked through this tube and the Mach number at the tube inlet. Answers: 0.0136 kg/s, 0.523
kPa. For an average friction factor of 0.014, determine the duct length from the inlet where the inlet velocity doubles. Also determine the pressure drop along that section of the duct. 12–128E Air flows through a 6-in-diameter, 50-ft-long adiabatic duct with inlet conditions of V1 ! 500 ft/s, T01 ! 650 R, and P1 ! 50 psia. For an average friction factor of 0.02, determine the velocity, temperature, and pressure at the exit of the duct. 12–129
Consider subsonic airflow through a 10-cmdiameter adiabatic duct with inlet conditions of T1 ! 330 K, P1 ! 180 kPa, and Ma1 ! 0.1. Taking the average friction factor to be 0.02, determine the duct length required to accelerate the flow to a Mach number of unity. Also, calculate the duct length at Mach number intervals of 0.1, and plot the duct length against the Mach number for 0.1 . Ma . 1. Discuss the results. 12–130
P0 ! 95 kPa T0 ! 290 K
D ! 1 cm
Vacuum pump
L ! 50 cm
Review Problems
FIGURE P12–123 12–124 Repeat Prob. 12–123 for a friction factor of 0.025 and a tube length of 1 m. 12–125 Air enters a 5-cm-diameter, 4-m-long adiabatic duct with inlet conditions of Ma1 ! 2.8, T1 ! 380 K, and P1 ! 80 kPa. It is observed that a normal shock occurs at a location 3 m from the inlet. Taking the average friction factor to be 0.007, determine the velocity, temperature, and pressure at the duct exit. Answers: 572 m/s, 813 K, 328 kPa P1 ! 80 kPa T1 ! 380 K Ma1 ! 2.8
Argon gas with k ! 1.667, cp ! 0.5203 kJ/kg · K, and R ! 0.2081 kJ/kg · K enters an 8cm-diameter adiabatic duct with V1 ! 70 m/s, T1 ! 520 K, and P1 ! 350 kPa. Taking the average friction factor to be 0.005 and letting the exit temperature T2 vary from 540 K to 400 K, evaluate the entropy change at intervals of 10 K, and plot the Fanno line on a T-s diagram.
Normal shock
L1 ! 3 m
FIGURE P12–125 12–126E Helium gas with k ! 1.667 enters a 6-in-diameter duct at Ma1 ! 0.2, P1 ! 60 psia, and T1 ! 600 R. For an average friction factor of 0.025, determine the maximum duct length that will not cause the mass flow rate of helium to be reduced. Answer: 291 ft 12–127 Air enters a 20-cm-diameter adiabatic duct with inlet conditions of V1 ! 150 m/s, T1 ! 500 K, and P1 ! 200
12–131 Air in an automobile tire is maintained at a pressure of 220 kPa (gage) in an environment where the atmospheric pressure is 94 kPa. The air in the tire is at the ambient temperature of 25°C. Now a 4-mm-diameter leak develops in the tire as a result of an accident. Assuming isentropic flow, determine the initial mass flow rate of air through the leak. Answer: 0.554 kg/min
12–132 The thrust developed by the engine of a Boeing 777 is about 380 kN. Assuming choked flow in the nozzles, determine the mass flow rate of air through the nozzle. Take the ambient conditions to be 295 K and 95 kPa. 12–133 A stationary temperature probe inserted into a duct where air is flowing at 250 m/s reads 85°C. What is the actual temperature of air? Answer: 53.9°C 12–134 Nitrogen enters a steady-flow heat exchanger at 150 kPa, 10°C, and 100 m/s, and it receives heat in the amount of 150 kJ/kg as it flows through it. The nitrogen leaves the heat exchanger at 100 kPa with a velocity of 200 m/s. Determine the stagnation pressure and temperature of the nitrogen at the inlet and exit states. 12–135 Derive an expression for the speed of sound based on van der Waals’ equation of state P ! RT(v $ b) $ a/v2. Using this relation, determine the speed of sound in carbon dioxide at 50°C and 200 kPa, and compare your result to that obtained by assuming ideal-gas behavior. The van der Waals constants for carbon dioxide are a ! 364.3 kPa · m6/kmol2 and b ! 0.0427 m3/kmol.
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12–136 Obtain Eq. 12–10 by starting with Eq. 12–9 and using the cyclic rule and the thermodynamic property relations cp cv &s &s and !a b . !a b T T &T P &T v
12–137 For ideal gases undergoing isentropic flows, obtain expressions for P/P*, T/T*, and r/r* as functions of k and Ma. 12–138 Using Eqs. 12–4, 12–13, and 12–14, verify that for the steady flow of ideal gases dT0/T ! dA/A " (1 $ Ma2) dV/V. Explain the effect of heating and area changes on the velocity of an ideal gas in steady flow for (a) subsonic flow and (b) supersonic flow. 12–139 A subsonic airplane is flying at a 3000-m altitude where the atmospheric conditions are 70.109 kPa and 268.65 K. A Pitot static probe measures the difference between the static and stagnation pressures to be 22 kPa. Calculate the speed of the airplane and the flight Mach number. Answers: 209 m/s, 0.637
# 12–140 Plot the mass flow parameter m 2RT0 /(AP0) versus the Mach number for k ! 1.2, 1.4, and 1.6 in the range of 0 . Ma . 1. 12–141 Helium enters a nozzle at 0.8 MPa, 500 K, and a velocity of 120 m/s. Assuming isentropic flow, determine the pressure and temperature of helium at a location where the velocity equals the speed of sound. What is the ratio of the area at this location to the entrance area? 12–142 Repeat Problem 12–141 assuming the entrance velocity is negligible. 12–143
Air at 0.9 MPa and 400 K enters a converging nozzle with a velocity of 180 m/s. The throat area is 10 cm2. Assuming isentropic flow, calculate and plot the mass flow rate through the nozzle, the exit velocity, the exit Mach number, and the exit pressure–stagnation pressure ratio versus the back pressure–stagnation pressure ratio for a back pressure range of 0.9 + Pb + 0.1 MPa. 12–144
Steam at 6.0 MPa and 700 K enters a converging nozzle with a negligible velocity. The nozzle throat area is 8 cm2. Assuming isentropic flow, plot the exit pressure, the exit velocity, and the mass flow rate through the nozzle versus the back pressure Pb for 6.0 + Pb + 3.0 MPa. Treat the steam as an ideal gas with k ! 1.3, cp ! 1.872 kJ/kg · K, and R ! 0.462 kJ/kg · K. 12–145 Find the expression for the ratio of the stagnation pressure after a shock wave to the static pressure before the shock wave as a function of k and the Mach number upstream of the shock wave Ma1. 12–146 Nitrogen enters a converging–diverging nozzle at 700 kPa and 300 K with a negligible velocity, and it experiences a normal shock at a location where the Mach number is Ma ! 3.0. Calculate the pressure, temperature, velocity,
Mach number, and stagnation pressure downstream of the shock. Compare these results to those of air undergoing a normal shock at the same conditions. 12–147 An aircraft flies with a Mach number Ma1 ! 0.8 at an altitude of 7000 m where the pressure is 41.1 kPa and the temperature is 242.7 K. The diffuser at the engine inlet has an exit Mach number of Ma2 ! 0.3. For a mass flow rate of 65 kg/s, determine the static pressure rise across the diffuser and the exit area. 12–148 Helium expands in a nozzle from 1 MPa, 500 K, and negligible velocity to 0.1 MPa. Calculate the throat and exit areas for a mass flow rate of 0.25 kg/s, assuming the nozzle is isentropic. Why must this nozzle be converging– diverging? Answers: 3.51 cm2, 5.84 cm2 12–149E Helium expands in a nozzle from 150 psia, 900 R, and negligible velocity to 15 psia. Calculate the throat and exit areas for a mass flow rate of 0.2 lbm/s, assuming the nozzle is isentropic. Why must this nozzle be converging– diverging? 12–150
Using the EES software and the relations in Table A–13, calculate the one-dimensional compressible flow functions for an ideal gas with k ! 1.667, and present your results by duplicating Table A–13. 12–151
Using the EES software and the relations in Table A–14, calculate the one-dimensional normal shock functions for an ideal gas with k ! 1.667, and present your results by duplicating Table A–14. 12–152 Consider an equimolar mixture of oxygen and nitrogen. Determine the critical temperature, pressure, and density for stagnation temperature and pressure of 800 K and 500 kPa. 12–153
Using EES (or other) software, determine the shape of a converging–diverging nozzle for air for a mass flow rate of 3 kg/s and inlet stagnation conditions of 1400 kPa and 200°C. Assume the flow is isentropic. Repeat the calculations for 50-kPa increments of pressure drops to an exit pressure of 100 kPa. Plot the nozzle to scale. Also, calculate and plot the Mach number along the nozzle. 12–154
Using EES (or other) software and the relations given in Table A–13, calculate the onedimensional isentropic compressible-flow functions by varying the upstream Mach number from 1 to 10 in increments of 0.5 for air with k ! 1.4. 12–155 12–156
Repeat Prob. 12–154 for methane with k ! 1.3.
Using EES (or other) software and the relations given in Table A–14, generate the one-dimensional normal shock functions by varying the upstream Mach number from 1 to 10 in increments of 0.5 for air with k ! 1.4.
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12–157
Repeat Prob. 12–156 for methane with k ! 1.3.
12–158 Air in a room at T0 ! 300 K and P0 ! 100 kPa is to be drawn by a vacuum pump through a 3-cm-diameter, 2m-long adiabatic tube equipped with a converging nozzle at the inlet. The flow in the nozzle section can be assumed to be isentropic. The static pressure is measured to be 97 kPa at the tube inlet and 55 kPa at the tube exit. Determine the mass flow rate of air through the duct, the air velocity at the duct exit, and the average friction factor for the duct. 12–159 Air enters a 4-cm-diameter adiabatic duct with inlet conditions of Ma1 ! 2.2, T1 ! 250 K, and P1 ! 80 kPa, and exits at a Mach number of Ma2 ! 1.8. Taking the average friction factor to be 0.03, determine the velocity, temperature, and pressure at the exit. 12–160 Air is cooled as it flows through a 20-cm-diameter duct. The inlet conditions are Ma1 ! 1.2, T01 ! 350 K, and P01 ! 240 kPa and the exit Mach number is Ma2 ! 2.0. Disregarding frictional effects, determine the rate of cooling of air. 12–161 Air is heated as it flows subsonically through a 10 cm # 10 cm square duct. The properties of air at the inlet are maintained at Ma1 ! 0.4, P1 ! 400 kPa, and T1 ! 360 K at all times. Disregarding frictional losses, determine the highest rate of heat transfer to the air in the duct without affecting the inlet conditions. Answer: 1958 kW Qmax P1 ! 400 kPa T1 ! 360 K Ma1 ! 0.4
enter a 10-cm-diameter adiabatic duct with inlet conditions of Ma1 ! 2, T1 ! 510 K, and P1 ! 180 kPa. If a normal shock occurs at a location 2 m from the inlet, determine the velocity, temperature, and pressure at the duct exit. Take the average friction factor of the duct to be 0.010. 12–166
Consider supersonic airflow through a 18-cmdiameter adiabatic duct with inlet conditions of T1 ! 530 K, P1 ! 80 kPa, and Ma1 ! 3. Taking the average friction factor to be 0.03, determine the duct length required to decelerate the flow to a Mach number of unity. Also, calculate the duct length at Mach number intervals of 0.25, and plot the duct length against the Mach number for 1 . Ma . 3. Discuss the results. 12–167
Air is flowing through a 6-cm-diameter adiabatic duct with inlet conditions of V1 ! 120 m/s, T1 ! 400 K, and P1 ! 100 kPa and an exit Mach number of Ma2 ! 1. To study the effect of duct length on the mass flow rate and the inlet velocity, the duct is now extended until its length is doubled while P1 and T1 are held constant. Taking the average friction factor to be 0.02, calculate the mass flow rate, and the inlet velocity, for various extension lengths, and plot them against the extension length. Discuss the results. 12–168 In compressible flow, velocity measurements with a Pitot probe can be grossly in error if relations developed for incompressible flow are used. Therefore, it is essential that compressible flow relations be used when evaluating flow velocity from Pitot probe measurements. Consider supersonic flow of air through a channel. A probe inserted into the flow causes a shock wave to occur upstream of the probe, and it measures the stagnation pressure and temperature to be 620 kPa and 340 K, respectively. If the static pressure upstream is 110 kPa, determine the flow velocity.
FIGURE P12–161 12–162
Repeat Prob. 12–161 for helium.
12–163 Air is accelerated as it is heated in a duct with negligible friction. Air enters at V1 ! 100 m/s, T1 ! 400 K, and P1 ! 35 kPa and the exits at a Mach number of Ma2 ! 0.8. Determine the heat transfer to the air, in kJ/kg. Also determine the maximum amount of heat transfer without reducing the mass flow rate of air. 12–164 Air at sonic conditions and at static temperature and pressure of 500 K and 420 kPa, respectively, is to be accelerated to a Mach number of 1.6 by cooling it as it flows through a channel with constant cross-sectional area. Disregarding frictional effects, determine the required heat transfer from the air, in kJ/kg. Answer: 69.8 kJ/kg 12–165 Combustion gases with an average specific heat ratio of k ! 1.33 and a gas constant of R ! 0.280 kJ/kg % K
P1 ! 110 kPa Shock wave
P02 ! 620 kPa T02 ! 340 K
FIGURE P12–168
Design and Essay Problems 12–169 Find out if there is a supersonic wind tunnel on your campus. If there is, obtain the dimensions of the wind tunnel and the temperatures and pressures as well as the Mach number at several locations during operation. For what typical experiments is the wind tunnel used?
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678 FLUID MECHANICS
12–170 Assuming you have a thermometer and a device to measure the speed of sound in a gas, explain how you can determine the mole fraction of helium in a mixture of helium gas and air. 12–171 Design a 1-m-long cylindrical wind tunnel whose diameter is 25 cm operating at a Mach number of 1.8. Atmospheric air enters the wind tunnel through a converging– diverging nozzle where it is accelerated to supersonic velocities. Air leaves the tunnel through a converging–diverging diffuser where it is decelerated to a very low velocity before entering the fan section. Disregard any irreversibilities. Spec-
ify the temperatures and pressures at several locations as well as the mass flow rate of air at steady-flow conditions. Why is it often necessary to dehumidify the air before it enters the wind tunnel?
P0 T0
Ma ! 1.8
FIGURE P12–171
D ! 25 cm
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CHAPTER
13
OPEN-CHANNEL FLOW
O
pen-channel flow implies flow in a channel open to the atmosphere, but flow in a conduit is also open-channel flow if the liquid does not fill the conduit completely, and thus there is a free surface. The pipe flow discussed in Chap. 8 involves closed conduits filled with a liquid or a gas. An open-channel flow, however, involves liquids only (typically water or wastewater) exposed to a gas (usually air, which is at atmospheric pressure). Flow in pipes is driven by a pressure difference, whereas flow in a channel is driven naturally by gravity. Water flow in a river, for example, is driven by the elevation difference between upstream and downstream. The flow rate in an open channel is established by the dynamic balance between gravity and friction. Inertia of the flowing liquid also becomes important in unsteady flow. The free surface coincides with the hydraulic grade line (HGL) and the pressure is constant along the free surface. But the height of the free surface from the channel bottom and thus all dimensions of the flow cross-section along the channel is not known a priori—it changes along with average flow velocity. The pressure in a channel varies hydrostatically in the vertical direction when the flow is steady and fully developed. In this chapter we present the basic principles of open-channel flows and the associated correlations for steady one-dimensional flow in channels of common cross sections. Detailed information can be obtained from several books written on the topic, some of which are listed in the references.
OBJECTIVES When you finish reading this chapter, you should be able to ■ Understand how flow in open channels differs from flow in pipes ■
■
■
Learn the different flow regimes in open channels and their characteristics Predict if hydraulic jumps are to occur during flow, and calculate the fraction of energy dissipated during hydraulic jumps Learn how flow rates in open channels are measured using sluice gates and weirs
679
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680 FLUID MECHANICS
13–1
FIGURE 13–1 Natural and human-made openchannel flows are characterized by a free surface open to the atmosphere. © Vol. 16/PhotoDisc.
2.0 1.5 1.0 0.5
FIGURE 13–2 Typical constant relative velocity curves in an open channel of trapezoidal cross section.
■
CLASSIFICATION OF OPEN-CHANNEL FLOWS
Open-channel flow refers to the flow of liquids in channels open to the atmosphere or in partially filled conduits and is characterized by the presence of a liquid–gas interface called the free surface (Fig. 13–1). Most natural flows encountered in practice, such as the flow of water in creeks, rivers, and floods, as well as the draining of rainwater off highways, parking lots, and roofs are open-channel flows. Human-made open-channel flow systems include irrigation systems, sewer lines, drainage ditches, and gutters, and the design of such systems is an important application area of engineering. In an open channel, the flow velocity is zero at the side and bottom surfaces because of the no-slip condition, and maximum at the midplane of the free surface (when there are significant secondary flows, as in the bends of noncircular channels, the maximum velocity occurs below the free surface somewhere within the top 25 percent of depth, as shown in Fig. 13–2). Furthermore, flow velocity also varies in the flow direction in most cases. Therefore, the velocity distribution (and thus flow) in open channels is, in general, three-dimensional. In engineering practice, however, the equations are written in terms of the average velocity at a cross section of the channel. Since the average velocity varies only with streamwise distance x, V is a onedimensional variable. The one-dimensionality makes it possible to solve significant real-world problems in a simple manner by hand calculations, and we restrict our consideration in this chapter to flows with one-dimensional average velocity. Despite its simplicity, the one-dimensional equations provide remarkably accurate results and are commonly used in practice. The no-slip condition on the channel walls gives rise to velocity gradients, and wall shear stress tw develops along the wetted surfaces. The wall shear stress varies along the wetted perimeter at a given cross section and offers resistance to flow. The magnitude of this resistance depends on the viscosity of the fluid as well as the velocity gradients at the wall surface. Open-channel flows are also classified as being steady or unsteady. A flow is said to be steady if there is no change with time at a given location. The representative quantity in open-channel flows is the flow depth (or alternately, the average velocity), which may vary along the channel. The flow is said to be steady if the flow depth does not vary with time at any given location along the channel (although it may vary from one location to another). Otherwise, the flow is unsteady. In this chapter we deal with steady flow only.
Uniform and Varied Flows
Flow in open channels is also classified as being uniform or nonuniform (also called varied), depending on how the flow depth y (the distance of the free surface from the bottom of the channel measured in the vertical direction) varies along the channel. The flow in a channel is said to be uniform if the flow depth (and thus the average velocity) remains constant. Otherwise, the flow is said to be nonuniform or varied, indicating that the flow depth varies with distance in the flow direction. Uniform flow conditions
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681 CHAPTER 13
are commonly encountered in practice in long straight sections of channels with constant slope and constant cross section. In open channels of constant slope and constant cross section, the liquid accelerates until the head loss due to frictional effects equals the elevation drop. The liquid at this point reaches its terminal velocity, and uniform flow is established. The flow remains uniform as long as the slope, cross section, and surface roughness of the channel remain unchanged. The flow depth in uniform flow is called the normal depth yn, which is an important characteristic parameter for open-channel flows (Fig. 13–3). The presence of an obstruction in the channel, such as a gate or a change in slope or cross section, causes the flow depth to vary, and thus the flow to become varied or nonuniform. Such varied flows are common in both natural and human-made open channels such as rivers, irrigation systems, and sewer lines. The varied flow is called rapidly varied flow (RVF) if the flow depth changes markedly over a relatively short distance in the flow direction (such as the flow of water past a partially open gate or over a falls), and gradually varied flow (GVF) if the flow depth changes gradually over a long distance along the channel. A gradually varied flow region typically occurs between rapidly varied and uniform flow regions, as shown in Fig. 13–4. In gradually varied flows, we can work with the one-dimensional average velocity just as we can with uniform flows. However, average velocity is not always the most useful or most appropriate parameter for rapidly varying flows. Therefore, the analysis of rapidly varied flows is rather complicated, especially when the flow is unsteady (such as the breaking of waves on the shore). For a known discharge rate, the flow height in a gradually varied flow region (i.e., the profile of the free surface) in a specified open channel can be determined in a step-by-step manner by starting the analysis at a cross section where the flow conditions are known, and evaluating head loss, elevation drop, and then the average velocity for each step.
Uniform flow y ! yn ! constant
V ! constant
Slope: S0 ! constant
FIGURE 13–3 For uniform flow in an open channel, the flow depth y and the average flow velocity V remain constant.
Laminar and Turbulent Flows in Channels
Like pipe flow, open-channel flow can be laminar, transitional, or turbulent, depending on the value of the Reynolds number expressed as Re !
UF
GVF
rVR h VR h ! m n
RVF
GVF
(13–1)
UF
FIGURE 13–4 Uniform flow (UF), gradually varied flow (GVF), and rapidly varied flow (RVF) in an open channel.
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682 FLUID MECHANICS I’ve known since grade school that radius is half of diameter. Now they tell me that hydraulic radius is one-fourth of hydraulic diameter!
?
?
?
Here V is the average liquid velocity, n is the kinematic viscosity, and Rh is the hydraulic radius defined as the ratio of the cross-sectional flow area Ac and the wetted perimeter p, Hydraulic radius:
Rh !
(m)
(13–2)
Considering that open channels come with rather irregular cross sections, the hydraulic radius serves as the characteristic dimension and brings uniformity to the treatment of open channels. Also, the Reynolds number is constant for the entire uniform flow section of an open channel. You might expect that the hydraulic radius would be defined as half the hydraulic diameter, but this is unfortunately not the case. Recall that the hydraulic diameter Dh for pipe flow is defined as Dh ! 4Ac/p so that the hydraulic diameter reduces to the pipe diameter for circular pipes. Then the relation between hydraulic radius and hydraulic diameter becomes Hydraulic diameter:
FIGURE 13–5 The relationship between the hydraulic radius and hydraulic diameter is not what you might expect.
Ac p
Dh !
4A c ! 4R h p
(13–3)
So, we see that the hydraulic radius is in fact one-fourth, rather than onehalf, of the hydraulic diameter (Fig. 13–5). Therefore, a Reynolds number based on the hydraulic radius is one-fourth of the Reynolds number based on hydraulic diameter as the characteristic dimension. So it will come as no surprise that the flow is laminar for Re " 2000 in pipe flow, but for Re " 500 in open-channel flow. Also, open-channel flow is usually turbulent for Re # 2500 and transitional for 500 " Re " 2500. Laminar flow is encountered when a thin layer of water (such as the rainwater draining off a road or parking lot) flows at a low velocity. The kinematic viscosity of water at 20°C is 1.00 $ 10%6 m2/s, and the average flow velocity in open channels is usually above 0.5 m/s. Also, the hydraulic radius is usually greater than 0.1 m. Therefore, the Reynolds number associated with water flow in open channels is typically above 50,000, and thus the flow is almost always turbulent. Note that the wetted perimeter includes the sides and the bottom of the channel in contact with the liquid—it does not include the free surface and the parts of the sides exposed to air. For example, the wetted perimeter and the cross-sectional flow area for a rectangular channel of height h and width b containing water of depth y are p ! b & 2y and Ac ! yb, respectively. Then, Rectangular channel:
Rh !
Ac yb y ! ! p b & 2y 1 & 2y/b
(13–4)
As another example, the hydraulic radius for the drainage of water of depth y off a parking lot of width b is (Fig. 13–6) Liquid layer of thickness y:
Rh !
Ac yb yb ! ! !y p b & 2y b
(13–5)
since b '' y. Therefore, the hydraulic radius for the flow of a liquid film over a large surface is simply the thickness of the liquid layer.
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683 CHAPTER 13
R
R u
y
y u b A y(b & y/tan u) Rh ! pc ! b & 2y/sin u
Ac ! R 2(u % sin u cos u) p ! 2Ru A u % sin u cos u Rh ! pc ! R 2u
(b) Trapezoidal channel
(a) Circular channel (u in rad)
yVb
y b b
A yb yb Rh ! pc ! b & 2y " b " y
A yb y Rh ! pc ! b & 2y ! 1 & 2y/b
(d) Liquid film of thickness y
FIGURE 13–6 Hydraulic radius relations for various open-channel geometries.
(c) Rectangular channel
13–2
■
FROUDE NUMBER AND WAVE SPEED
Open-channel flow is also classified as tranquil, critical, or rapid, depending on the value of the dimensionless Froude number discussed in Chap. 7 and defined as Fr !
Froude number:
V 2gL c
!
V 2gy
(13–6)
where g is the gravitational acceleration, V is the average liquid velocity at a cross section, and Lc is the characteristic length, which is taken to be the flow depth y for wide rectangular channels. The Froude number is an important parameter that governs the character of flow in open channels. The flow is classified as Fr ( 1
Subcritical or tranquil flow
Fr ! 1
Critical flow
Fr ' 1
Supercritical or rapid flow
(13–7)
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684 FLUID MECHANICS
Compressible Flow
Open-Channel Flow
Ma ! V/c
Fr ! V/c0
Ma ( 1 Subsonic Ma ! 1 Sonic Ma ' 1 Supersonic
Fr ( 1 Subcritical Fr ! 1 Critical Fr ' 1 Supercritical
V ! speed of flow c ! #kRT kRT ! speed of sound (ideal gas) c0 ! #gy gy ! speed of wave (liquid)
FIGURE 13–7 Analogy between the Mach number in compressible flow and the Froude number in open-channel flow.
yc
y
Subcritical flow: y ' yc
This resembles the classification of compressible flow with respect to the Mach number: subsonic for Ma ( 1, sonic for Ma ! 1, and supersonic for Ma ' 1 (Fig. 13–7). Indeed, the denominator of the Froude number has the dimensions of velocity, and it represents the speed c0 at which a small disturbance travels in still liquid, as shown later in this section. Therefore, in analogy to the Mach number, the Froude number is expressed as the ratio of the flow speed to the wave speed, Fr ! V/c0, just as the Mach number is expressed as the ratio of the flow speed to sound speed, Ma ! V/c. The Froude number can also be thought of as the square root of the ratio of inertia (or dynamic) force to gravity force (or weight). This can be demonstrated by multiplying both the numerator and the denominator of the square of the Froude number V 2/gLc by rA, where r is density and A is a representative area, which gives Fr 2 !
1 V 2 rA 2( 2 rV 2A) Inertia force ) ! mg gL c rA Gravity force
y
Supercritical flow: y ( yc
FIGURE 13–8 Definitions of subcritical flow and supercritical flow in terms of critical depth.
(13–8)
Here Lc A represents volume, rLc A is the mass of this fluid volume, and mg is the weight. The numerator is twice the inertial force 12 rV 2A, which can be thought of as the dynamic pressure 12 rV 2 times the cross-sectional area, A. Therefore, the flow in an open channel is dominated by inertial forces in rapid flow and by gravity forces in tranquil flow. It follows that at low flow velocities (Fr ( 1), a small disturbance travels upstream (with a velocity c0 % V relative to a stationary observer) and affects the upstream conditions. This is called tranquil or subcritical flow. But at high flow velocities (Fr ' 1), a small disturbance cannot travel upstream (in fact, the wave is washed downstream at a velocity of V % c0 relative to a stationary observer) and thus the upstream conditions cannot be influenced by the downstream conditions. This is called rapid or supercritical flow, and the flow in this case is controlled by the upstream conditions. Therefore, a surface wave travels upstream when Fr ( 1, is swept downstream when Fr ' 1, and appears frozen on the surface when Fr ! 1. Also, the surface wave speed increases with flow depth y, and thus a surface disturbance propagates much faster in deep channels than it does in shallow ones. Consider the flow of a liquid in an open rectangular channel of cross. with a volume flow rate of V . When the flow is critical, sectional area Ac 1gy Fr ! 1 and the average flow velocity is V ! , where yc is the critic # cal depth. Noting that V ! A cV ! A c 1gy c, the critical depth can be expressed as Critical depth (general):
yc
# V2 yc ! 2 gA c
(13–9)
For a rectangular channel of width b we have Ac ! byc, and the critical depth relation reduces to Critical depth (rectangular):
# V 2 1/3 y c ! a 2b gb
(13–10)
The liquid depth is y ' yc for subcritical flow and y ( yc for supercritical flow (Fig. 13–8). As in compressible flow, a liquid can accelerate from subcritical to supercritical flow. Of course, it can also decelerate from supercritical to subcriti-
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685 CHAPTER 13
cal flow, but it can do so by undergoing a shock. The shock in this case is called a hydraulic jump, which corresponds to a normal shock in compressible flow. Therefore, the analogy between open-channel flow and compressible flow is remarkable.
Speed of Surface Waves
We are all familiar with the waves forming on the free surfaces of oceans, lakes, rivers, and even swimming pools. The surface waves can be very high, like the ones we see on the oceans, or barely noticeable. Some are smooth; some break on the surface. A basic understanding of wave motion is necessary for the study of certain aspects of open-channel flow, and here we present a brief description. A detailed treatment of wave motion can be found in numerous books written on the subject. An important parameter in the study of open-channel flow is the wave speed c0 , which is the speed at which a surface disturbance travels through a liquid. Consider a long, wide channel that initially contains a still liquid of height y. One end of the channel is moved with speed dV, generating a surface wave of height dy propagating at a speed of c0 into the still liquid, as shown in Fig. 13–9a. Now consider a control volume that encloses the wave front and moves with it, as shown in Fig. 13–9b. To an observer traveling with the wave front, the liquid to the right appears to be moving toward the wave front with speed c0 and the liquid to the left appears to be moving away from the wave front with speed c0 % dV. Of course the observer would think the control volume that encloses the wave front (and herself or himself) is stationary, and he or she would be witnessing a steady-flow process. . . The steady-flow mass balance m1 ! m2 (or the continuity relation) for this control volume of width b can be expressed as rc0 yb ! r(c0 % dV)(y & dy)b
→
dV ! c0
dy y & dy
(13–11)
We make the following assumptions: (1) the velocity is nearly constant across the channel and thus the momentum flux correction factors (b1 and b2) are one, (2) the distance across the wave is short and thus friction at the bottom surface and air drag at the top are negligible, (3) the dynamic effects are negligible and thus the pressure in the liquid varies hydrostatically; in terms of gage pressure, P1, avg ! rgh1, avg ! rg(y/ 2) and P2, avg ! rgh2, avg . . ! rg(y & dy)/2, (4) the mass flow rate is constant with m1 ! m2 ! rc0yb, and (5) there are no external forces or body forces and thus the only forces acting on the control volume in the horizontal x-direction are the pressure → # → # → forces. Then, the momentum equation a F ! a bmV % a bmV in the out
in
x-direction becomes a balance between hydrostatic pressure forces and momentum transfer, # # P2, avg A 2 % P1, avg A 1 ! m (%V2) % m (%V1)
(13–12)
Note that both the inlet and the outlet average velocities are negative since they are in the negative x-direction. Substituting, rg(y & dy)2b rgy 2b % ! rc0yb(%c0 & dV) % rc0yb(%c0) 2 2
(13–13)
Moving position
dy c0 dV y
Still liquid
Moving wavefront (a) Generation and propagation of a wave
Control dy volume
y
c0%dV c0 rg(y & dy) (2)
(1)
rgy
(b) Control volume relative to an observer traveling with the wave, with gage pressure distributions shown
FIGURE 13–9 The generation and analysis of a wave in an open channel.
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686 FLUID MECHANICS
or,
Sluice gate
ga1 & Hydraulic jump Subcritical flow
Supercritical flow
dy b dy ! c0 dV 2y
(13–14)
Combining the momentum and continuity relations and rearranging give c 20 ! gya1 &
Subcritical flow
dy dy b a1 & b y 2y
(13–15)
Therefore, the wave speed c0 is proportional to the wave height dy. For infinitesimal surface waves, dy (( y and thus FIGURE 13–10 Supercritical flow through a sluice gate.
(a)
(b)
FIGURE 13–11 A hydraulic jump can be observed on a dinner plate when (a) it is right-sideup, but not when (b) it is upside down. Photos by Abel Po-Ya Chuang. Used by permission.
Infinitesimal surface waves:
c0 ! 2gy
(13–16)
Therefore, the speed of infinitesimal surface waves is proportional to the square root of liquid depth. Again note that this analysis is valid only for shallow water bodies, such as those encountered in open channels. Otherwise, the wave speed is independent of liquid depth for deep bodies of water, such as the oceans. The wave speed can also be determined by using the energy balance relation instead of the momentum equation together with the continuity relation. Note that the waves eventually die out because of the viscous effects that are neglected in the analysis. Also, for flow in channels of non-rectangular cross-section, the hydraulic depth defined as yh ! Ac/Lt where Lt is the top width of the flow section should be used in the calculation of Froude number in place of the flow depth y. For a half-full circular channel, for example, the hydraulic depth is yh ! (pR2/ 2)/ 2R ! pR/4. We know from experience that when a rock is thrown into a lake, the concentric waves that form propagate evenly in all directions and vanish after some distance. But when the rock is thrown into a river, the upstream side of the wave moves upstream if the flow is tranquil or subcritical (V ( c0 ), moves downstream if the flow is rapid or supercritical (V ' c0 ), and remains stationary at the location where it is formed if the flow is critical (V ! c0 ). You may be wondering why we pay so much attention to flow being subcritical or supercritical. The reason is that the character of the flow is strongly influenced by this phenomenon. For example, a rock at the riverbed may cause the water level at that location to rise or to drop, depending on whether the flow is subcritical or supercritical. Also, the liquid level drops gradually in the flow direction in subcritical flow, but a sudden rise in liquid level, called a hydraulic jump, may occur in supercritical flow (Fr ' 1) as the flow decelerates to subcritical (Fr ( 1) velocities. This phenomenon can occur downstream of a sluice gate as shown in Fig. 13–10. The liquid approaches the gate with a subcritical velocity, but the upstream liquid level is sufficiently high to accelerate the liquid to a supercritical level as it passes through the gate (just like a gas flowing in a converging–diverging nozzle). But if the downstream section of the channel is not sufficiently sloped down, it cannot maintain this supercritical velocity, and the liquid jumps up to a higher level with a larger cross-sectional area, and thus to a lower subcritical velocity. Finally, the flow in rivers, canals, and irrigation systems is typically subcritical. But the flow past sluice gates and spillways is typically supercritical. You can create a beautiful hydraulic jump the next time you wash dishes (Fig. 13–11). Let the water from the faucet hit the middle of a dinner plate. As
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687 CHAPTER 13
the water spreads out radially, its depth decreases and the flow is supercritical. Eventually, a hydraulic jump occurs, which you can see as a sudden increase in water depth. Try it!
13–3
■
SPECIFIC ENERGY
Consider the flow of a liquid in a channel at a cross section where the flow depth is y, the average flow velocity is V, and the elevation of the bottom of the channel at that location relative to some reference datum is z. For simplicity, we ignore the variation of liquid speed over the cross section and assume the speed to be V everywhere. The total mechanical energy of this liquid in the channel in terms of heads is expressed as (Fig. 13–12) 2
H!z&
2g Es
y
2
V V P & !z&y& rg 2g 2g
(13–17)
where z is the elevation head, P/rg ! y is the gage pressure head, and V 2/2g is the velocity or dynamic head. The total energy as expressed in Eq. 13–17 is not a realistic representation of the true energy of a flowing fluid since the choice of the reference datum and thus the value of the elevation head z is rather arbitrary. The intrinsic energy of a fluid at a cross section can be represented more realistically if the reference datum is taken to be the bottom of the channel so that z ! 0 there. Then the total mechanical energy of a fluid in terms of heads becomes the sum of the pressure and dynamic heads. The sum of the pressure and dynamic heads of a liquid in an open channel is called the specific energy Es and is expressed as (Bakhmeteff, 1932) Es ! y &
Energy line V2
V2 2g
z
Reference datum
FIGURE 13–12 The specific energy Es of a liquid in an open channel is the total mechanical energy (expressed as a head) relative to the bottom of the channel.
(13–18)
as shown in Fig. 13–12. Consider flow in an. open channel of constant width b. Noting that the volume flow rate is V ! AcV ! ybV, the average flow velocity can be expressed as V!
# V yb
(13–19)
Substituting into Eq. 13–18, the specific energy can be expressed as Es ! y &
# V2 2gb 2y 2
y (13–20)
This equation is very instructive as it shows the variation of the specific energy with flow depth. During steady flow in an open channel the flow rate . is constant, and a plot of Es versus y for constant V and b is given in Fig. 13–13. We observe the following from this figure: • The distance from a point on the vertical y-axis to the curve represents the specific energy at that y-value. The part between the Es ! y line and the curve corresponds to dynamic head (or kinetic energy) of the liquid, and the remaining part to pressure head (or flow energy). • The specific energy tends to infinity as y → 0 (due to the velocity approaching infinity), and it becomes equal to flow depth y for large values of y (due to the velocity and thus the kinetic energy becoming very
. V ! constant
Es ! y
V2 2g
Subcritical flow, Fr ( 1
y yc Fr ! 1 Es, min
Critical depth
Supercritical flow, Fr ' 1 Es
FIGURE 13–13 Variation of specific energy Es with depth y for a specified flow rate.
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688 FLUID MECHANICS
small). The specific energy reaches a minimum value Es, min at some intermediate point, called the critical point, characterized by the critical depth yc and critical velocity Vc. The minimum specific energy is also called the critical energy. • There is a minimum. specific energy Es, min required to support the . specified flow rate V. Therefore, Es cannot be below Es, min for a given V. • A horizontal line intersects the specific energy curve at one point only, and thus a fixed value of flow depth corresponds to a fixed value of specific. energy. This is expected since the velocity has a fixed value when V, b, and y are specified. However, for Es ' Es, min, a vertical line intersects the curve at two points, indicating that a flow can have two different depths (and thus two different velocities) corresponding to a fixed value of specific energy. These two depths are called alternate depths. For flow through a sluice gate with negligible frictional losses (and thus Es ! constant), the upper depth corresponds to the upstream flow, and the lower depth to the downstream flow. • A small change in specific energy near the critical point causes a large difference between alternate depths and may cause violent fluctuations in flow level. Therefore, operation near the critical point should be avoided in the design of open channels. The value of the minimum specific energy and the critical depth at which it occurs can be determined by . differentiating Es from Eq. 13–20 with respect to y for constant b and V, and setting the derivative equal to zero: # # dE s d V2 V2 ! ay & b !1% 2 3!0 dy dy 2gb 2y 2 gb y
(13–21)
# V 2 1/3 y c ! a 2b gb
(13–22)
Vc ! 2gy c
(13–23)
Solving for y, which is the critical flow depth yc, gives
. The flow rate at the critical point can be expressed as V ! ycbVc. Substituting, the critical velocity is determined to be which is the wave speed. The Froude number at this point is Fr !
V 2gy
!
Vc 2gy c
!1
(13–24)
indicating that the point of minimum specific energy is indeed the critical point, and the flow becomes critical when the specific energy reaches its minimum value. It follows that the flow is subcritical at lower flow velocities and thus higher flow depths (the upper arm of the curve), supercritical at higher velocities and thus lower flow depths (the lower arm of the curve), and critical at the critical point (the point of minimum specific energy). Noting that Vc ! 1gy c, the minimum (or critical) specific energy can be expressed in terms of the critical depth alone as E s, min ! y c &
V 2c gy c 3 ! yc & ! yc 2g 2g 2
(13–25)
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689 CHAPTER 13
In uniform flow, the flow depth and the flow velocity, and thus the specific energy, remain constant since E s ! y & V 2/2g. The head loss is made up by the decline in elevation (the channel is sloped downward in the flow direction). In nonuniform flow, however, the specific energy may increase or decrease, depending on the slope of the channel and the frictional losses. If the decline in elevation across a flow section is more than the head loss in that section, for example, the specific energy increases by an amount equal to the difference between elevation drop and head loss. The specific energy concept becomes a particularly useful tool when studying varied flows.
EXAMPLE 13–1
Character of Flow and Alternate Depth
0.2 m3/s
Water is flowing steadily in a 0.4-m-wide rectangular open channel at a rate of 0.2 m3/s (Fig. 13–14). If the flow depth is 0.15 m, determine the flow velocity and if the flow is subcritical or supercritical. Also determine the alternate flow depth if the character of flow were to change.
SOLUTION Water flow in a rectangular open channel is considered. The character of flow, the flow velocity, and the alternate depth are to be determined. Assumptions The specific energy is constant. Analysis The average flow velocity is determined from
# # 0.2 m3/s V V ! 3.33 m/s V! ! ! A c yb (0.15 m)(0.4 m)
The critical depth for this flow is
yc ! a
# 1/3 (0.2 m3/s)2 V 2 1/3 b !a b ! 0.294 m 2 2 2 gb (9.81 m/s )(0.4 m)
Therefore, the flow is supercritical since the actual flow depth is y ! 0.15 m, and y ( yc. Another way to determine the character of flow is to calculate the Froude number,
Fr !
3.33 m/s V ! ! 2.75 1gy 2(9.81 m/s2)(0.15 m)
Again the flow is supercritical since Fr ' 1. The specific energy for the given conditions is
# (0.2 m3/s)2 V2 E s1 ! y 1 & ! (0.15 m) & ! 0.7163 m 2gb 2y 21 2(9.81 m/s2)(0.4 m)2(0.15 m)2
Then the alternate depth is determined from Es1 ! Es2 to be
# V2 E s2 ! y 2 & 2gb 2y 22
→
0.7163 m ! y 2 &
(0.2 m3/s)2 2(9.81 m/s2)(0.4 m)2y 22
Solving for y2 gives the alternate depth to be y2 ! 0.69 m. Therefore, if the character of flow were to change from supercritical to subcritical while holding the specific energy constant, the flow depth would rise from 0.15 to 0.69 m. Discussion Note that if water underwent a hydraulic jump at constant specific energy (the frictional losses being equal to the drop in elevation), the flow depth would rise to 0.69 m, assuming of course that the side walls of the channel are high enough.
0.15 m 0.4 m
FIGURE 13–14 Schematic for Example 13–1.
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690 FLUID MECHANICS Energy line
H!z&y& V2 2g
V2 2g
13–4
■
CONTINUITY AND ENERGY EQUATIONS
Open-channel flows involve liquids whose densities are nearly constant, and thus the one-dimensional steady-flow conservation of mass or continuity equation can be expressed as # V ! A cV ! constant
a y
V
A y%a
That is, the product of the flow cross section and the average flow velocity remains constant throughout the channel. The continuity equation between two sections along the channel is expressed as Continuity equation:
z Reference datum
FIGURE 13–15 The total energy of a liquid flowing in an open channel.
(13–26)
A c1V1 ! A c2V2
(13–27)
which is identical to the steady-flow continuity equation for liquid flow in a pipe. Note that both the flow cross section and the average flow velocity may vary during flow, but, as stated, their product remains constant. To determine the total energy of a liquid flowing in an open channel relative to a reference datum, as shown in Fig. 13–15, consider a point A in the liquid at a distance a from the free surface (and thus a distance y % a from the channel bottom). Noting that the elevation, pressure (hydrostatic pressure relative to the free surface), and velocity at point A are zA ! z & (y % a), PA ! rga, and VA ! V, respectively, the total energy of the liquid in terms of heads is HA ! z A &
rga V 2 PA V 2A V2 & & ! z & (y % a) & !z&y& rg 2g rg 2g 2g
(13–28)
which is independent of the location of the point A at a cross section. Therefore, the total mechanical energy of a liquid at any cross section of an open channel can be expressed in terms of heads as H!z&y&
V2 2g
(13–29)
where y is the flow depth, z is the elevation of the channel bottom, and V is the average flow velocity. Then the one-dimensional energy equation for open-channel flow between an upstream section 1 and a downstream section 2 can be written as Energy equation:
z1 & y1 &
V 21 V 22 ! z2 & y2 & & hL 2g 2g
(13–30)
The head loss hL due to frictional effects is expressed as in pipe flow as hL ! f
L V2 L V2 !f D h 2g R h 8g
(13–31)
where f is the average friction factor and L is the length of channel between sections 1 and 2. The relation Dh ! 4Rh should be observed when using the hydraulic radius instead of the hydraulic diameter. Flow in open channels is gravity driven, and thus a typical channel is slightly sloped down. The slope of the bottom of the channel is expressed as S 0 ! tan a !
z1 % z2 z1 % z2 ! x2 % x1 L
(13–32)
where a is the angle the channel bottom makes with the horizontal. In general, the bottom slope S0 is very small, and thus the channel bottom is
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691 CHAPTER 13
nearly horizontal. Therefore, L " x2 % x1, where x is the distance in the horizontal direction. Also, the flow depth y, which is measured in the vertical direction, can be taken to be the depth normal to the channel bottom with negligible error. If the channel bottom is straight so that the bottom slope is constant, the vertical drop between sections 1 and 2 can be expressed as z 1 % z 2 ! S 0L. Then the energy equation (Eq. 13–30) becomes Energy equation:
y1 &
V 21 V 22 & S 0L ! y 2 & & hL 2g 2g
(13–33)
This equation has the advantage that it is independent of a reference datum for elevation. In the design of open-channel systems, the bottom slope is selected such that it provides adequate elevation drop to overcome the frictional head loss and thus to maintain flow at the desired rate. Therefore, there is a close connection between the head loss and the bottom slope, and it makes sense to express the head loss as a slope (or the tangent of an angle). This is done by defining a friction slope as Friction slope:
Sf !
hL L
V12 2g V1
(13–34)
Then the energy equation can be written as Energy equation:
y1 &
V 21 2g
! y2 &
V 22 2g
& (S f % S 0)L
y1 (1)
(13–35)
z1
z
Note that the friction slope is equal to the bottom slope when the head loss is equal to the elevation drop. That is, Sf ! S0 when hL ! z1 % z2. Figure 13–16 also shows the energy line, which is a distance z & y & V 2/2g (total mechanical energy of the liquid expressed as a head) above the horizontal reference datum. The energy line is typically sloped down like the channel itself as a result of frictional losses, the vertical drop being equal to the head loss hL. Note that if there were no head loss, the energy line would be horizontal even when the channel is not. The elevation and velocity heads (z & y and V 2/2g) would then be able to convert to each other during flow in this case, but their sum would remain constant.
hL
Energy line
L
V 22 2g V2
(2) a Slope: S0 ! constant z 2 Horizontal reference datum
x1
y2
x2
x
FIGURE 13–16 The total energy of a liquid at two sections of an open channel.
y1
V1 ! V2 ! V0 y1 ! y2 ! yn
y2
(1)
13–5
■
UNIFORM FLOW IN CHANNELS
We mentioned in Sec. 13–1 that flow in a channel is called. uniform flow if the flow depth (and thus the average flow velocity since V ! AcV ! constant in steady flow) remains constant. Uniform flow conditions are commonly encountered in practice in long straight runs of channels with constant slope, constant cross section, and constant surface lining. In the design of open channels, it is very desirable to have uniform flow in the majority of the system since this means having a channel of constant height, which is easier to design and build. The flow depth in uniform flow is called the normal depth yn, and the average flow velocity is called the uniform-flow velocity V0. The flow remains uniform as long as the slope, cross section, and surface roughness of the channel remain unchanged (Fig. 13–17). When the bottom slope is
z
a (2) z1 Slope: S0 ! tan a ! constant z 2 x2 % x1 ! Lcosa ! L x1
x2
x
Head loss ! elevation loss hL ! z 1 % z 2 ! S0 L
FIGURE 13–17 In uniform flow, the flow depth y, the average flow velocity V, and the bottom slope S0 remain constant, and the head loss equals the elevation loss, hL ! z1 % z2 ! Sf L ! S0L.
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increased, the flow velocity increases and the flow depth decreases. Therefore, a new uniform flow will be established with a new (lower) flow depth. The opposite occurs if the bottom slope is decreased. During flow in open channels of constant slope S0, constant cross section Ac , and constant surface friction factor f, the terminal velocity is reached and thus uniform flow is established when the head loss equals the elevation drop. Therefore, hL ! f
L V2 D h 2g
or
S 0L ! f
L V 20 R h 8g
(13–36)
since hL ! S0L in uniform flow and Dh ! 4Rh. Solving the second relation for V0, the uniform-flow velocity and the flow rate are determined to be V0 ! C2S 0R h
and
# V ! CA c 2S 0R h
(13–37)
where C ! 28g/f
(13–38)
is called the Chezy coefficient. The Eqs. 13–37 and the coefficient C are named in honor of the French engineer Antoine Chezy (1718–1798), who first proposed a similar relationship in about 1769. The Chezy coefficient is a dimensional quantity, and its value ranges from about 30 m1/2/s for small channels with rough surfaces to 90 m1/2/s for large channels with smooth surfaces (or, 60 ft1/2/s to 160 ft1/2/s in English units). The Chezy coefficient can be determined in a straightforward manner from Eq. 13–38 by first determining the friction factor f as done for pipe flow in Chap. 8 from the Moody chart or the Colebrook equation for the fully rough turbulent limit (Re → *), f ! [2.0 log(14.8R h /e)]%2
(13–39)
Here e is the mean surface roughness. Note that open-channel flow is typically turbulent, and the flow is fully developed by the time uniform flow is established. Therefore, it is reasonable to use the friction factor relation for fully developed turbulent flow. Also, at large Reynolds numbers, the friction factor curves corresponding to specified relative roughness are nearly horizontal, and thus the friction factor is independent of the Reynolds number. The flow in that region is called fully rough turbulent flow. Since the introduction of the Chezy equations, considerable effort has been devoted by numerous investigators to the development of simpler empirical relations for the average velocity and flow rate. The most widely used equation was developed independently by the Frenchman PhilippeGaspard Gauckler (1826–1905) in 1868 and the Irishman Robert Manning (1816–1897) in 1889. Both Gauckler and Manning made recommendations that the constant in the Chezy equation be expressed as a C ! R 1/6 n h
(13–40)
where n is called the Manning coefficient, whose value depends on the roughness of the channel surfaces. Substituting into Eqs. 13–37 gives the following empirical relations known as the Manning equations (also referred to as Gauckler–Manning equations since they were first proposed by Philippe-Gaspard Gauckler) for the uniform-flow velocity and the flow rate,
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693 CHAPTER 13
Uniform flow:
a S 1/2 V0 ! R 2/3 n h 0
# a 1/2 V ! A c R 2/3 h S0 n
and
(13–41)
The factor a is a dimensional constant whose value in SI units is a ! 1 m1/3/s. Noting that 1 m ! 3.2808 ft, its value in English units is a ! 1 m1/3/s ! (3.2808 ft)1/3/s ! 1.486 ft1/3/s
(13–42)
Note that the bottom slope S0 and the Manning coefficient n are dimensionless quantities, and Eqs. 13–41 give the velocity in m/s and the flow rate in m3/s in SI units when Rh is expressed in m. (The corresponding units in English units are ft/s and ft3/s when Rh is expressed in ft.) Experimentally determined values of n are given in Table 13–1 for numerous natural and artificial channels. More extensive tables are available in the literature. Note that the value of n varies from 0.010 for a glass channel to 0.150 for a floodplain laden with trees (15 times that of a glass channel). There is considerable uncertainty in the value of n, especially in natural channels, as you would expect, since no two channels are exactly alike. The scatter can be 20 percent or more. Also, the coefficient n depends on the size and shape of the channel as well as the surface roughness.
Critical Uniform Flow
Flow through an open channel becomes critical flow when the Froude number Fr ! 1 and thus the flow speed equals the wave speed Vc ! 1gy c, flow depth, defined previously (Eq. 13–9). When the where yc is the critical . volume flow rate V, the channel slope S0, and the Manning coefficient n are known, the normal flow depth yn can be determined from the Manning equation (Eq. 13–41). However, since Ac and Rh are both functions of yn, the equation often ends up being implicit in yn and requires a numerical (or trial and error) approach to solve. If yn ! yc, the flow is uniform critical flow, and bottom slope S0 equals the critical slope. Sc in this case. When flow depth yn is known instead of the flow rate V, the flow rate can be determined from the Manning equation and the critical flow depth from Eq. 13–9. Again the flow is critical only if yn ! yc. . During uniform critical flow, S0 ! Sc and yn ! yc. Replacing V and S0 in # the Manning equation by V ! A c 1gy c and Sc, respectively, and solving for Sc gives the following general relation for the critical slope, Critical slope (general):
Sc !
gn 2y c a 2R 4/3 h
(13–43)
For film flow or flow in a wide rectangular channel with b '' yc, Eq. 13–43 simplifies to Critical slope (b '' yc):
Sc !
gn 2 a 2y 1/3 c
(13–44)
This equation gives the slope necessary to maintain a critical flow of depth yc in a wide rectangular channel having a Manning coefficient of n.
Superposition Method for Nonuniform Perimeters
The surface roughness and thus the Manning coefficient for most natural and some human-made channels vary along the wetted perimeter and even along the channel. A river, for example, may have a stony bottom for its
TA B L E 1 3 – 1 Mean values of the Manning coefficient n for water flow in open channels* From Chow (1959).
Wall Material A. Artificially lined channels Glass Brass Steel, smooth Steel, painted Steel, riveted Cast iron Concrete, finished Concrete, unfinished Wood, planed Wood, unplaned Clay tile Brickwork Asphalt Corrugated metal Rubble masonry B. Excavated earth channels Clean Gravelly Weedy Stony, cobbles C. Natural channels Clean and straight Sluggish with deep pools Major rivers Mountain streams D. Floodplains Pasture, farmland Light brush Heavy brush Trees
n 0.010 0.011 0.012 0.014 0.015 0.013 0.012 0.014 0.012 0.013 0.014 0.015 0.016 0.022 0.025 0.022 0.025 0.030 0.035 0.030 0.040 0.035 0.050 0.035 0.050 0.075 0.150
* The uncertainty in n can be + 20 percent or more.
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694 FLUID MECHANICS
regular bed but a surface covered with bushes for its extended floodplain. There are several methods for solving such problems, either by finding an effective Manning coefficient n for the entire channel cross section, or by considering the channel in subsections and applying the superposition principle. For example, a channel cross section can be divided into N subsections, each with its own uniform Manning coefficient and flow rate. When determining the perimeter of a section, only the wetted portion of the boundary for that section is considered, and the imaginary boundaries are ignored. The flow rate through the channel is the sum of the flow rates through all the sections. EXAMPLE 13–2
y ! 0.52 m u ! 60° b ! 0.8 m
FIGURE 13–18 Schematic for Example 13–2.
Flow Rate in an Open Channel in Uniform Flow
Water is flowing in a weedy excavated earth channel of trapezoidal cross section with a bottom width of 0.8 m, trapezoid angle of 60°, and a bottom slope angle of 0.3°, as shown in Fig. 13–18. If the flow depth is measured to be 0.52 m, determine the flow rate of water through the channel. What would your answer be if the bottom angle were 1°?
SOLUTION Water is flowing in a weedy trapezoidal channel of given dimensions. The flow rate corresponding to a measured value of flow depth is to be determined. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties The Manning coefficient for an open channel with weedy surfaces is n ! 0.030. Analysis The cross-sectional area, perimeter, and hydraulic radius of the channel are y 0.52 m b ! (0.52 m)a0.8 m & b ! 0.5721 m2 tan u tan 602y 2 $ 0.52 m ! 0.8 m & ! 2.001 m p!b& sin u sin 60A c 0.5721 m2 ! 0.2859 m Rh ! ! p 2.991 m A c ! yab &
The bottom slope of the channel is
S 0 ! tan a ! tan 0.3- ! 0.005236 Then the flow rate through the channel is determined from the Manning equation to be
# a 1 m1/3,s 1,2 V ! A cR 2,3 (0.5721 m2)(0.2859 m)2/3(0.005236)1,2 ! 0.60 m3/s h S0 ! n 0.030 The flow rate for a bottom angle of 1° can be determined by using S0 . ! tan a ! tan 1° ! 0.01746 in the last relation. It gives V ! 1.1 m3/s. Discussion Note that the flow rate is a strong function of the bottom angle. Also, there is considerable uncertainty in the value of the Manning coefficient, and thus in the flow rate calculated. A 10 percent uncertainty in n results in a 10 percent uncertainty in the flow rate. Final answers are therefore given to only two significant digits.
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695 CHAPTER 13
EXAMPLE 13–3
The Height of a Rectangular Channel
Water is to be transported in an unfinished-concrete rectangular channel with a bottom width of 4 ft at a rate of 51 ft3/s. The terrain is such that the channel bottom drops 2 ft per 1000 ft length. Determine the minimum height of the channel under uniform-flow conditions (Fig. 13–19). What would your answer be if the bottom drop is just 1 ft per 1000 ft length?
SOLUTION Water is flowing in an unfinished-concrete rectangular channel with a specified bottom width. The minimum channel height corresponding to a specified flow rate is to be determined. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties The Manning coefficient for an open channel with unfinishedconcrete surfaces is n ! 0.014. Analysis The cross-sectional area, perimeter, and hydraulic radius of the channel are A c ! by ! (4 ft)y
p ! b & 2y ! (4 ft) & 2y
Rh !
Ac 4y ! p 4 & 2y
The bottom slope of the channel is S0 ! 2/1000 ! 0.002. Using the Manning equation, the flow rate through the channel can be expressed as
# a 1,2 V ! A cR 2,3 h S0 n 51 ft3/s !
2,3 4y 1.486 ft1/3,s (4y ft2)a ftb (0.002)1,2 0.014 4 & 2y
which is a nonlinear equation in y. Using an equation solver such as EES or an itirative approach, the flow depth is determined to be
y ! 2.5 ft If the bottom drop were just 1 ft per 1000 ft length, the bottom slope would be S0 ! 0.001, and the flow depth would be y ! 3.3 ft. Discussion Note that y is the flow depth, and thus this is the minimum value for the channel height. Also, there is considerable uncertainty in the value of the Manning coefficient n, and this should be considered when deciding the height of the channel to be built.
EXAMPLE 13–4
Channels with Nonuniform Roughness
Water flows in a channel whose bottom slope is 0.003 and whose cross section is shown in Fig. 13–20. The dimensions and the Manning coefficients for the surfaces of different subsections are also given on the figure. Determine the flow rate through the channel and the effective Manning coefficient for the channel.
SOLUTION Water is flowing through a channel with nonuniform surface properties. The flow rate and the effective Manning coefficient are to be determined.
. V ! 51 ft3/s
y
b ! 4 ft
FIGURE 13–19 Schematic for Example 13–3.
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696 FLUID MECHANICS 6m
8m
1
2
Clean natural channel n 1 ! 0.030
Light brush n 2 ! 0.050
2m
3m
FIGURE 13–20 Schematic for Example 13–4.
s
Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The Manning coefficients do not vary along the channel. Analysis The channel involves two parts with different roughnesses, and thus it is appropriate to divide the channel into two subsections as indicated in Fig. 13–20. The flow rate for each subsection can be determined from the Manning equation, and the total flow rate can be determined by adding them up. The side length of the triangular channel is s ! 132 & 32 ! 4.243 m. Then the flow area, perimeter, and hydraulic radius for each subsection and the entire channel become
Subsection 1: A c1 ! 21 m2
p1 ! 10.486 m
R h1 !
A c1 21 m2 ! ! 2.00 m p1 10.486 m
Subsection 2: A c2 ! 16 m2
p2 ! 10 m
R h2 !
A c2 16 m2 ! ! 1.60 m p2 10 m
Entire channel: A c ! 37 m2
p ! 20.486 m
Rh !
Ac 37 m2 ! ! 1.806 m p 20.486 m
Using the Manning equation for each subsection, the total flow rate through the channel is determined to be
# # # a a 1,2 V ! V 1 & V 2 ! A c1R 2,3 A R 2,3S 1,2 h1 S 0 & n1 n2 c2 h2 0 (21 m2)(2 m)2/3 (16 m2)(1.60 m)2/3 & d(0.003)1,2 0.030 0.050
! (1 m1/3,s) c
! 84.8 m3/s ! 85 m3/s
Knowing the total flow rate, the effective Manning coefficient for the entire channel can be determined from the Manning equation to be
n eff !
1,2 aA cR 2,3 (1 m1/3,s)(37 m2)(1.806 m)2/3(0.003)1,2 h S0 ! ! 0.035 # 84.8 m3,s V
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697 CHAPTER 13
Discussion The effective Manning coefficient neff of the channel turns out to lie between the two n values, as expected. The weighted average of the Manning coefficient of the channel is navg ! (n1p1 & n2p2)/p ! 0.040, which is quite different than neff. Therefore, using a weighted average Manning coefficient for the entire channel may be tempting, but it would not be so accurate.
13–6
■
BEST HYDRAULIC CROSS SECTIONS
Open-channel systems are usually designed to transport a liquid to a location at a lower elevation at a specified rate under the influence of gravity at the lowest possible cost. Noting that no energy input is required, the cost of an open-channel system consists primarily of the initial construction cost, which is proportional to the physical size of the system. Therefore, for a given channel length, the perimeter of the channel is representative of the system cost, and it should be kept to a minimum in order to minimize the size and thus the cost of the system. From another perspective, resistance to flow is due to wall shear stress tw and the wall area, which is equivalent to the wetted perimeter per unit channel length. Therefore, for a given flow cross-sectional area Ac, the smaller the wetted perimeter p, the smaller the resistance force, and thus the larger the average velocity and the flow rate. From yet another perspective, for a specified channel geometry with a specified bottom slope S0 and surface lining (and thus the roughness coefficient n), 1,2 the flow velocity is given by the Manning formula as V ! aR 2,3 h S 0 ,n. Therefore, the flow velocity is proportional to the hydraulic radius, and the hydraulic radius must be maximized (and thus the perimeter must be minimized since Rh ! Ac/p) in order to maximize the average flow velocity or the flow rate per unit cross-sectional area. Thus we conclude the following: The best hydraulic cross section for an open channel is the one with the maximum hydraulic radius or, equivalently, the one with the minimum wetted perimeter for a specified cross section.
The shape with the minimal perimeter per unit area is a circle. Therefore, on the basis of minimum flow resistance, the best cross section for an open channel is a semicircular one (Fig. 13–21). However, it is usually cheaper to construct an open channel with straight sides (such as channels with trapezoidal or rectangular cross sections) instead of semicircular ones, and the general shape of the channel may be specified a priori. Thus it makes sense to analyze each geometric shape separately for the best cross section. As a motivational example, consider a rectangular channel of finished concrete (n ! 0.012) of width b and flow depth y with a bottom slope of 1° (Fig. 13–22). To determine the effects. of the aspect ratio y/b on the hydraulic radius R. h and the flow rate V per unit cross-sectional area (Ac ! 1 m2), Rh and V are evaluated from the Manning formula. The results are tabulated in Table 13–2 and plotted in Fig. 13–23 for aspect ratios. from 0.1 to 5. We observe from this table and the plot that the flow rate V increases as the flow aspect ratio y/b is increased, reaches a maximum at y/b ! 0.5,
R y
FIGURE 13–21 The best hydraulic cross section for an open channel is a semicircular one since it has the minimum wetted perimeter for a specified cross section, and thus the minimum flow resistance.
y
b
FIGURE 13–22 A rectangular open channel of width b and flow depth y. For a given cross-sectional area, the highest flow rate occurs when y ! b/2.
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TA B L E 1 3 – 2
. Variation of the hydraulic radius Rh and the flow rate V with aspect ratio y/b for a rectangular channel with Ac ! 1 m2, S0 ! tan 1°, and n ! 0.012 Aspect Ratio y/b
Channel Width b, m
Flow Depth y, m
Perimeter p, m
Hydraulic Radius Rh, m
Flow.Rate V, m3/s
0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.5 2.0 3.0 4.0 5.0
3.162 2.236 1.826 1.581 1.414 1.291 1.195 1.118 1.054 1.000 0.816 0.707 0.577 0.500 0.447
0.316 0.447 0.548 0.632 0.707 0.775 0.837 0.894 0.949 1.000 1.225 1.414 1.732 2.000 2.236
3.795 3.130 2.921 2.846 2.828 2.840 2.869 2.907 2.951 3.000 3.266 3.536 4.041 4.500 4.919
0.264 0.319 0.342 0.351 0.354 0.352 0.349 0.344 0.339 0.333 0.306 0.283 0.247 0.222 0.203
4.53 5.14 5.39 5.48 5.50 5.49 5.45 5.41 5.35 5.29 5.00 4.74 4.34 4.04 3.81
1
2
5.75
. Flow rate V, m3/s
5.35
4.95
4.55
4.15
FIGURE 13–23 Variation of the flow rate in a rectangular channel with aspect ratio r ! y/b for Ac ! 1 m2 and S0 ! tan 1-.
3.75
0
3
4
5
Aspect ratio r ! y/b
. and then starts to decrease (the numerical values for V can also be interpreted as the flow velocities in m/s since Ac ! 1 m2). We see the same trend for the hydraulic radius, but the opposite trend for the wetted perimeter p. These results confirm that the best cross section for a given shape is the one with the maximum hydraulic radius, or equivalently, the one with the minimum perimeter.
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Rectangular Channels
Consider liquid flow in an open channel of rectangular cross section of width b and flow depth y. The cross-sectional area and the wetted perimeter at a flow section are and
A c ! yb
p ! b & 2y
(13–45)
Solving the first relation of Eq. 13–45 for b and substituting it into the second relation give p!
Ac & 2y y
(13–46)
Now we apply the criterion that the best hydraulic cross section for an open channel is the one with the minimum wetted perimeter for a given cross section. Taking the derivative of p with respect to y while holding Ac constant gives Ac by dp b !% 2&2!% 2&2!% &2 y dy y y
(13–47)
Setting dp/dy ! 0 and solving for y, the criterion for the best hydraulic cross section is determined to be Best hydraulic cross section (rectangular channel):
y!
b 2
(13–48)
Therefore, a rectangular open channel should be designed such that the liquid height is half the channel width to minimize flow resistance or to maximize the flow rate for a given cross-sectional area. This also minimizes the perimeter and thus the construction costs. This result confirms the finding from Table 13–2 that y ! b/2 gives the best cross section.
Trapezoidal Channels
Now consider liquid flow in an open channel of trapezoidal cross section of bottom width b, flow depth y, and trapezoid angle u measured from the horizontal, as shown in Fig. 13–24. The cross-sectional area and the wetted perimeter at a flow section are A c ! ab &
y by tan u
and
p!b&
2y sin u
u (13–49) b
Solving the first relation of Eq. 13–49 for b and substituting it into the second relation give p!
Ac y 2y % & y tan u sin u
(13–50)
Taking the derivative of p with respect to y while holding Ac and u constant gives Ac b & y,tan u dp 1 1 2 2 % !% 2% & !% & y dy tan u sin u tan u sin u y
(13–51)
Setting dp/dy ! 0 and solving for y, the criterion for the best hydraulic cross section for any specified trapezoid angle u is determined to be Best hydraulic cross section (trapezoidal channel):
y!
b sin u 2(1 % cos u)
y s
(13–52)
Rh !
Ac y(b & y/tan u) ! p b & 2y/sin u
FIGURE 13–24 Parameters for a trapezoidal channel.
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700 FLUID MECHANICS
For the special case of u ! 90° (a rectangular channel), this relation reduces to y ! b/2, as expected. The hydraulic radius Rh for a trapezoidal channel can be expressed as Rh !
A c y(b $ y%tan u) y(b sin u $ y cos u) ! ! p b $ 2y%sin u b sin u $ 2y
(13–53)
Rearranging Eq. 13–52 as bsinu ! 2y(1 " cos u), substituting into Eq. 13–53 and simplifying, the hydraulic radius for a trapezoidal channel with the best cross section becomes Hydraulic radius for the best cross section:
Rh !
y 2
(13–54)
Therefore, the hydraulic radius is half the flow depth for trapezoidal channels with the best cross section regardless of the trapezoid angle u. Similarly, the trapezoid angle for the best hydraulic cross section is determined by taking the derivative of p (Eq. 13–50) with respect to u while holding Ac and y constant, setting dp/du ! 0, and solving the resulting equation for u. This gives u ! 60#
Best trapezoid angle:
(13–55)
Substituting the best trapezoid angle u ! 60° into the best hydraulic cross section relation y ! b sin u/(2 " 2 cos u) gives Best flow depth for u ! 60°:
y!
23 b 2
(13–56)
Then the length of the side edge of the flow section and the flow area become s! y! b
p ! 3b
3 b 2 60° b
Rh !
y 3 ! b 2 4
y b23%2 ! !b sin 60# 23%2
Ac !
3 3 2 b 4
FIGURE 13–25 The best cross section for trapezoidal channels is half of a hexagon.
A c ! ab $
(13–57) (13–58)
y b23%2 323 2 b y ! ab $ b (b23%2) ! b tan u tan 60# 4
(13–59)
since tan 60# ! 23. Therefore, the best cross section for trapezoidal channels is half of a hexagon (Fig. 13–25). This is not surprising since a hexagon closely approximates a circle, and a half-hexagon has the least perimeter per unit cross-sectional area of all trapezoidal channels. Best hydraulic cross sections for other channel shapes can be determined in a similar manner. For example, the best hydraulic cross section for a circular channel of diameter D can be shown to be y ! D/2. EXAMPLE 13–5
Best Cross Section of an Open Channel
Water is to be transported at a rate of 2 m3/s in uniform flow in an open channel whose surfaces are asphalt lined. The bottom slope is 0.001. Determine the dimensions of the best cross section if the shape of the channel is (a) rectangular and (b) trapezoidal (Fig. 13–26).
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701 CHAPTER 13
SOLUTION Water is to be transported in an open channel at a specified rate. The best channel dimensions are to be determined for rectangular and trapezoidal shapes. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties The Manning coefficient for an open channel with asphalt lining is n ! 0.016. Analysis (a) The best cross section for a rectangular channel occurs when the flow height is half the channel width, y ! b/2. Then the cross-sectional area, perimeter, and hydraulic radius of the channel are A c ! by !
b2 2
p ! b & 2y ! 2b
Rh !
b b y! ! 2 2
b
Ac b ! p 4
Substituting into the Manning equation,
# a 1,2 V ! A cR 2,3 h S0 n
→
# 2(0.016)(2 m3/s)4 2/3 3/8 2nV 4 2/3 3/8 b! a b !a b a2S 0 (1 m1/3,s)20.001
which gives b ! 1.84 m. Therefore, Ac ! 1.70 m2, p ! 3.68 m, and the dimensions of the best rectangular channel are
b ! 1.84 m
and
y ! 0.92 m
(b) The best cross section for a trapezoidal channel occurs when the trapezoid angle is 60° and flow height is y ! b13,2. Then,
A c ! y(b & b cos u) ! 0.523b 2(1 & cos 60-) ! 0.7523b 2 p ! 3b
y 23 Rh ! ! b 2 4
Substituting into the Manning equation,
# a 1,2 V ! A c R 2,3 h S0 n
→
3/8 (0.016)(2 m3/s) b! a b 2/3 1/3 0.7523a 23/4b (1 m ,s) 20.001
which yields b ! 1.12 m. Therefore, Ac ! 1.64 m2, p ! 3.37 m, and the dimensions of the best trapezoidal channel are
b ! 1.12 m
y ! 0.973 m
and
u ! 60-
Discussion Note that the trapezoidal cross section is better since it has a smaller perimeter (3.37 m versus 3.68 m) and thus lower construction cost.
13–7
■
GRADUALLY VARIED FLOW
To this point we considered uniform flow during which the flow depth y and the flow velocity V remain constant. In this section we consider gradually varied flow (GVF), which is a form of steady nonuniform flow characterized by gradual variations in flow depth and velocity (small slopes and no abrupt changes) and a free surface that always remains smooth (no discontinuities or zigzags). Flows that involve rapid changes in flow depth and velocity, called rapidly varied flows (RVF), are considered in Section 13–8. A change in the bottom slope or cross section of a channel or an obstruction
y! b
3 b 2 60° b
FIGURE 13–26 Schematic for Example 13–5.
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702 FLUID MECHANICS
Horizontal
Friction slope Sf
V 2 Energy line, H (V & dV)2 2g 2g
V
dhL
V & dV
y
y & dy
z, H
zb
Bottom slope S0
zb & dzb
dx x
Horizontal reference datum
x & dx x
FIGURE 13–27 Variation of properties over a differential flow section in an open channel under conditions of gradually varied flow (GVF).
in the path of flow may cause the uniform flow in a channel to become gradually or rapidly varied flow. Rapidly varied flows occur over a short section of the channel with relatively small surface area, and thus frictional losses associated with wall shear are negligible. Head losses in RVF are highly localized and are due to intense agitation and turbulence. Losses in GVF, on the other hand, are primarily due to frictional effects along the channel and can be determined from the Manning formula. In gradually varied flow, the flow depth and velocity vary slowly, and the free surface is stable. This makes it possible to formulate the variation of flow depth along the channel on the basis of the conservation of mass and energy principles and to obtain relations for the profile of the free surface. In uniform flow, the slope of the energy line is equal to the slope of the bottom surface. Therefore, the friction slope equals the bottom slope, Sf ! S0. In gradually varied flow, however, these slopes are different. Consider steady flow in a rectangular open channel of width b, and assume any variation in the bottom slope and water depth to be rather gradual. We again write the equations in terms of average velocity V and assume the pressure distribution to be hydrostatic. From Eq. 13–17, the total head of the liquid at any cross section is H ! zb & y & V 2/2g, where zb is the vertical distance of the bottom surface from the reference datum. Differentiating H with respect to x gives dz b dy V dV dH d V2 ! az b & y & b ! & & dx dx 2g dx dx g dx
(13–60)
But H is the total energy of the liquid and thus dH/dx is the slope of the energy line (negative quantity), which is equal to the negative of the friction slope, as shown in Fig. 13–27. Also, dzb /dx is the negative of the bottom slope. Therefore, dhL dH !% ! %S f dx dx
dz b ! %S 0 dx
and
(13–61)
Substituting Eq. 13–61 into Eq. 13–60 gives S0 % Sf !
dy V dV & dx g dx
(13–62)
. The continuity equation for steady flow in a rectangular channel is V ! ybV ! constant. Differentiating with respect to x gives 0 ! bV
dy dV & yb dx dx
→
dV V dy !% y dx dx
(13–63)
Substituting Eq. 13–63 into Eq. 13–62 and noting that V, 1gy is the Froude number, S0 % Sf !
dy V 2 dy dy dy % ! % Fr 2 dx gy dx dx dx
(13–64)
Solving for dy/dx gives the desired relation for the rate of change of flow depth (or the surface profile) in gradually varied flow in an open channel, dy S 0 % S f ! dx 1 % Fr 2
(13–65)
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703 CHAPTER 13
which is analogous to the variation of flow area as a function of the Mach number in compressible flow. This relation is derived for a rectangular channel, but it is also valid for channels of other constant cross sections provided that the Froude number is expressed accordingly. An analytical or numerical solution of this differential equation gives the flow depth y as a function of x for a given set of parameters, and the function y(x) is the surface profile. The general trend of flow depth—whether it increases, decreases, or remains constant along the channel—depends on the sign of dy/dx, which depends on the signs of the numerator and the denominator of Eq. 13–65. The Froude number is always positive and so is the friction slope Sf (except for the idealized case of flow with negligible frictional effects for which both hL and Sf are zero). The bottom slope S0 is positive for downward-sloping sections (typically the case), zero for horizontal sections, and negative for upward-sloping sections of a channel (adverse flow). The flow depth increases when dy/dx ' 0, decreases when dy/dx ( 0, and remains constant (and thus the free surface is parallel to the channel bottom, as in uniform flow) when dy/dx ! 0 and thus S0 ! Sf. For specified values of S0 and Sf , the term dy/dx may be positive or negative, depending on whether the Froude number is less than or greater than 1. Therefore, the flow behavior is opposite in subcritical and supercritical flows. For S0 % Sf ' 0, for example, the flow depth increases in the flow direction in subcritical flow, but it decreases in supercritical flow. The determination of the sign of the denominator 1 % Fr2 is easy: it is positive for subcritical flow (Fr ( 1), and negative for supercritical flow (Fr ' 1). But the sign of the numerator depends on the relative magnitudes of S0 and Sf. Note that the friction slope Sf is always positive, and its value is equal to the channel slope S0 in uniform flow, y ! yn. Noting that head loss increases with increasing velocity, and that the velocity is inversely proportional to flow depth for a given flow rate, Sf ' S0 and thus S0 % Sf ( 0 when y ( yn, and Sf ( S0 and thus S0 % Sf ' 0 when y ' yn. The numerator S0 % Sf is always negative for horizontal (S0 ! 0) and upward-sloping (S0 ( 0) channels, and thus the flow depth decreases in the flow direction during subcritical flows in such channels.
Liquid Surface Profiles in Open Channels, y(x)
Open-channel systems are designed and built on the basis of the projected flow depths along the channel. Therefore, it is important to be able to predict the flow depth for a specified flow rate and specified channel geometry. A plot of flow depths gives the surface profile of the flow. The general characteristics of surface profiles for gradually varied flow depend on the bottom slope and flow depth relative to the critical and normal depths. A typical open channel involves various sections of different bottom slopes S0 and different flow regimes, and thus various sections of different surface profiles. For example, the general shape of the surface profile in a downward-sloping section of a channel is different than that in an upwardsloping section. Likewise, the profile in subcritical flow is different than the profile in supercritical flow. Unlike uniform flow that does not involve inertia
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704 FLUID MECHANICS A S H
Steep
Adverse
C Horizontal Critical
M
Mild
FIGURE 13–28 Designation of the letters S, C, M, H, and A for liquid surface profiles for different types of slopes.
1 Free surface in uniform flow 2 Free surface in critical flow
yn
3
yc y Channel bottom
FIGURE 13–29 Designation of the numbers 1, 2, and 3 for liquid surface profiles based on the value of the flow depth relative to the normal and critical depths.
forces, gradually varied flow involves acceleration and deceleration of liquid, and the surface profile reflects the dynamic balance between liquid weight, shear force, and inertial effects. Each surface profile is identified by a letter that indicates the slope of the channel and by a number that indicates flow depth relative to the critical depth yc and normal depth yn. The slope of the channel can be mild (M), critical (C), steep (S), horizontal (H), or adverse (A). The channel slope is said to be mild if yn ' yc, steep if yn ( yc, critical if yn ! yc, horizontal if S0 ! 0 (zero bottom slope), and adverse if S0 ( 0 (negative slope). Note that a liquid flows uphill in an open channel that has an adverse slope (Fig. 13–28). The classification of a channel section depends on the flow rate and the channel cross section as well as the slope of the channel bottom. A channel section that is classified to have a mild slope for one flow can have a steep slope for another flow, and even a critical slope for a third flow. Therefore, we need to calculate the critical depth yc and the normal depth yn before we can assess the slope. The number designation indicates the initial position of the liquid surface for a given channel slope relative to the surface levels in critical and uniform flows, as shown in Fig. 13–29. A surface profile is designated by 1 if the flow depth is above both critical and normal depths (y ' yc and y ' yn), by 2 if the flow depth is between the two (yn ' y ' yc or yn ( y ( yc), and by 3 if the flow depth is below both the critical and normal depths (y ( yc and y ( yn). Therefore, three different profiles are possible for a specified type of channel slope. But for channels with zero or adverse slopes, type 1 flow cannot exist since the flow can never be uniform in horizontal and upward channels, and thus normal depth is not defined. Also, type 2 flow does not exist for channels with critical slope since normal and critical depths are identical in this case. The five classes of slopes and the three types of initial positions discussed give a total of 12 distinct configurations for surface profiles in GVF, all tabulated and sketched in Table 13–3. The Froude number is also given for each case, with Fr ' 1 for y ( yc, as well as the sign of the slope dy/dx of the surface profile determined from Eq. 13–65, dy/dx ! (S0 % Sf)/(1 % Fr2). Note that dy/dx ' 0, and thus the flow depth increases in the flow direction when both S0 % Sf and 1 % Fr2 are positive or negative. Otherwise dy/dx ( 0 and the flow depth decreases. In type 1 flows, the flow depth increases in the flow direction and the surface profile approaches the horizontal plane asymptotically. In type 2 flows, the flow depth decreases and the surface profile approaches the lower of yc or yn. In type 3 flows, the flow depth increases and the surface profile approaches the lower of yc or yn. These trends in surface profiles continue as long as there is no change in bottom slope or roughness. Consider the first case in Table 13–3 designated M1 (mild channel slope and y ' yn ' yc ). The flow is subcritical since y ' yc and thus Fr ( 1 and 1 % Fr2 ' 0. Also, Sf ( S0 and thus S0 % Sf ' 0 since y ' yn, and thus the flow velocity is less than the velocity in normal flow. Therefore, the slope of the surface profile dy/dx ! (S0 % Sf)/(1 % Fr2) ' 0, and the flow depth y increases in the flow direction. But as y increases, the flow velocity decreases, and thus Sf and Fr approach zero. Consequently, dy/dx
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705 CHAPTER 13
TABLE 13–3 Classification of surface profiles in gradually varied flow Channel Slope Mild (M)
yc ! yn S0 ! Sc
M1 M2
M3
Profile Notation
Flow Depth
Froude Number
M1
y " yn
Fr ! 1
M2
yc ! y ! yn
Fr ! 1
M3
y ! yc
Fr " 1
Profile Slope
Surface Profile Starting point
dy "0 dx dy !0 dx dy "0 dx
yn Normal depth
Horizontal M1 Surface profile y(x) M2
yc Critical depth M3
Steep (S) yc " yn S0 ! Sc
S1
S1 S2
y " yc
Fr ! 1
S2
yn ! y ! yc
Fr " 1
S3
y ! yn
Fr " 1
S1
Channel bottom, S 0 ! Sc
dy "0 dx dy !0 dx dy "0 dx
Horizontal S1
yc
yn
S2 S3
S3
Channel bottom, S 0 " Sc
Critical (C) C1
yc # yn S0 ! Sc
C1
y " yc
Fr ! 1
C3
y ! yc
Fr " 1
C3
dy "0 dx dy "0 dx
Horizontal C1
yc # yn C3
Channel bottom, S 0 # Sc
Horizontal (H) yn → $ S0 # 0 H2 H3
H2
y " yc
Fr ! 1
H3
y ! yc
Fr " 1
H2
dy !0 dx dy "0 dx
H2 yc H3
Adverse (A) S0 ! 0 yn: does not exist A2 A2
A2
y " yc
Fr ! 1
A3
y ! yc
Fr " 1
dy !0 dx dy "0 dx
Channel bottom, S 0 # 0
A2 yc A3
A3 Channel bottom, S 0 ! 0
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706 FLUID MECHANICS
approaches S0 and the rate of increase in flow depth becomes equal to the channel slope. This requires the surface profile to become horizontal at large y. Then we conclude that the M1 surface profile first rises in the flow direction and then tends to a horizontal asymptote. As y → yc in subcritical flow (such as M2, H2, and A2), we have Fr → 1 and 1 % Fr2 → 0, and thus the slope dy/dx tending to negative infinity. But as y → yc in supercritical flow (such as M3, H3, and A3), we have Fr → 1 and 1 % Fr2 → 0, and thus the slope dy/dx, which is a positive quantity, tending to infinity. That is, the free surface rises almost vertically and the flow depth increases very rapidly. This cannot be sustained physically, and the free surface breaks down. The result is a hydraulic jump. The onedimensional assumption is no longer applicable when this happens.
Some Representative Surface Profiles
A typical open-channel system involves several sections of different slopes, with connections called transitions, and thus the overall surface profile of flow is a continuous profile made up of the individual profiles described earlier. Some representative surface profiles commonly encountered in open channels, including some composite profiles, are given in Fig. 13–30. For each case, the change in surface profile is caused by a change in channel geometry such as an abrupt change in slope or an obstruction in the flow such as a sluice gate. More composite profiles can be found in specialized books listed in the references. A point on a surface profile represents the flow height at that point that satisfies the mass, momentum, and energy conservation relations. Note that dy/dx (( 1 and S0 (( 1 in gradually varied flow, and the slopes of both the channels and the surface profiles in these sketches are highly exaggerated for better visualization. Many channels and surface profiles would appear nearly horizontal if drawn to scale. Figure 13–30a shows the surface profile for gradually varied flow in a channel with mild slope and a sluice gate. The subcritical upstream flow (note that the flow is subcritical since the slope is mild) slows down as it approaches the gate (such as a river approaching a dam) and the liquid level rises. The flow past the gate is supercritical (since the height of the opening is less than the critical depth). Therefore, the surface profile is M1 before the gate and M3 after the gate prior to the hydraulic jump. A section of an open channel may have a negative slope and involve uphill flow, as shown in Fig. 13–30b. Flow with an adverse slope cannot be maintained unless the inertia forces overcome the gravity and viscous forces that oppose the fluid motion. Therefore, an uphill channel section must be followed by a downhill section or a free outfall. For subcritical flow with an adverse slope approaching a sluice gate, the flow depth decreases as the gate is approached, yielding an A2 profile. Flow past the gate is typically supercritical, yielding an A3 profile prior to the hydraulic jump. The open-channel section in Fig. 13–30c involves a slope change from steep to less steep. The flow velocity in the less steep part is lower (a smaller elevation drop to drive the flow), and thus the flow depth is higher when uniform flow is established again. Noting that uniform flow with steep slope must be supercritical (y ( yc), the flow depth increases from the initial to the new uniform level smoothly through an S3 profile.
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707 CHAPTER 13
yn1
Uniform flow
Hydraulic jump
M1
Uniform flow
yn2
M3
yc yn2
Mild (a) Flow through a sluice gate in an open channel with mild slope Hydraulic jump A2
A2
A3 yc Adverse (b) Flow through a sluice gate in an open channel with adverse slope and free outfall yc yn1
Unif
orm
flow
S3
Uniform flow
yn2
y ! yn2
Steep
Less steep (c) Uniform supercritical flow changing from steep to less steep slope
yn1
Uniform flow
yc
M2 S2 yn2
Uniform flow
Mild
Hydraulic jump H3
H2
Steep
yc
Free outfall
Horizontal (d) Uniform subcritical flow changing from mild to steep to horizontal slope with free outfall
Figure 13–30d shows a composite surface profile for an open channel that involves various flow sections. Initially the slope is mild, and the flow is uniform and subcritical. Then the slope changes to steep, and the flow becomes supercritical when uniform flow is established. The critical depth occurs at the break in grade. The change of slope is accompanied by a smooth decrease in flow depth through an M2 profile at the end of the mild
FIGURE 13–30 Some common surface profiles encountered in open-channel flow. All flows are from left to right.
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708 FLUID MECHANICS
section, and through an S2 profile at the beginning of the steep section. In the horizontal section, the flow depth increases first smoothly through an H3 profile, and then rapidly during a hydraulic jump. The flow depth then decreases through an H2 profile as the liquid accelerates toward the end of the channel to a free outfall. The flow becomes critical before reaching the end of the channel, and the outfall controls the upstream flow past the hydraulic jump. The outfalling flow stream is supercritical. Note that uniform flow cannot be established in a horizontal channel since the gravity force has no component in the flow direction, and the flow is inertia-driven.
Numerical Solution of Surface Profile
The prediction of the surface profile y(x) is an important part of the design of open-channel systems. A good starting point for the determination of the surface profile is the identification of the points along the channel, called the control points, at which the flow depth can be calculated from a knowledge of flow rate. For example, the flow depth at a section of a rectangular channel where critical . flow occurs, called the critical point, can be determined from yc ! (V 2/gb2)1/3. The normal depth yn, which is the flow depth reached when uniform flow is established, also serves as a control point. Once flow depths at control points are available, the surface profile upstream or downstream can be determined usually by numerical integration of the nonlinear differential equation dy S 0 % S f ! dx 1 % Fr 2
(13–66)
The friction slope Sf is determined from the uniform-flow conditions, and the Froude number from a relation appropriate for the channel cross section. EXAMPLE 13–6
Classification of Channel Slope
Water is flowing uniformly in a rectangular open channel with unfinishedconcrete surfaces. The channel width is 6 m, the flow depth is 2 m, and the bottom slope is 0.004. Determine if the channel should be classified as mild, critical, or steep for this flow (Fig. 13–31). b ! y!2m 2
b!6m S0 ! 0.004
FIGURE 13–31 Schematic for Example 13–6.
SOLUTION Water is flowing in an open channel uniformly. It is to be determined whether the channel slope is mild, critical, or steep for this flow. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties The Manning coefficient for an open channel with unfinishedconcrete surfaces is n ! 0.014. Analysis The cross-sectional area, perimeter, and hydraulic radius are A c ! yb ! (2 m)(6 m) ! 12 m2 p ! b & 2y ! 6 m & 2(2 m) ! 10 m Rh !
A c 12 m2 ! ! 1.2 m p 10 m
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709 CHAPTER 13
The flow rate is determined from the Manning equation to be
# 1 m1/3,s a 1,2 (12 m2)(1.2 m)2/3(0.004)1,2 ! 61.2 m3/s V ! A c R 2,3 h S0 ! n 0.014 Noting that the flow is uniform, the specified flow rate is the normal depth and thus y ! yn ! 2 m. The critical depth for this flow is
yc !
# (61.2 m3,s)2 V2 ! ! 2.2 m gA 2c (9.81 m/s2)(12 m2)
This channel at these flow conditions is classified as steep since yn ( yc , and the flow is supercritical. Discussion If the flow depth were greater than 2.2 m, the channel slope would be said to be mild. Therefore, the bottom slope alone is not sufficient to classify a downhill channel as being mild, critical, or steep.
13–8
■
RAPIDLY VARIED FLOW AND HYDRAULIC JUMP
Recall that flow in open channels is called rapidly varied flow (RVF) if the flow depth changes markedly over a relatively short distance in the flow direction (Fig. 13–32). Such flows occur in sluice gates, broad- or sharpcrested weirs, waterfalls, and the transition sections of channels for expansion and contraction. A change in the cross section of the channel is an important reason for the occurrence of rapidly varied flow. But some rapidly varied flows, such as flow through a sluice gate, occur even in regions where the channel cross section is constant. Rapidly varied flows are typically complicated by the fact that they may involve significant multidimensional and transient effects, backflows, and flow separation. Therefore, rapidly varied flows are usually studied experimentally or numerically. But despite these complexities, it is still possible to analyze some rapidly varied flows using the one-dimensional flow approximation with reasonable accuracy. The flow in steep channels can be supercritical, and the flow can change to subcritical if the channel can no longer sustain supercritical flow due to a reduced slope of the channel or increased frictional effects. Any such change from supercritical to subcritical flow occurs through a hydraulic jump. A hydraulic jump involves considerable mixing and agitation, and thus a significant amount of mechanical energy dissipation. Consider steady flow through a control volume that encloses the hydraulic jump, as shown in Fig. 13–33. To make a simple analysis possible, we make the following assumptions: 1. The velocity is nearly constant across the channel at sections 1 and 2, and therefore the momentum-flux correction factors are b1 ! b2 " 1. 2. The pressure in the liquid varies hydrostatically, and we consider gage pressure only since atmospheric pressure acts on all surfaces and its effect cancels out.
FIGURE 13–32 Rapidly varied flow occurs when there is a sudden change in flow, such as an abrupt change in cross section. hL
Control volume
Energy line V2
y2
V1
y1 rgy1 (1)
(2) rgy2
x
y
Es2 ! y2 &
V 22 2g
2 Subcritical
1
Supercritical Es1
Es
FIGURE 13–33 Schematic and flow depth-specific energy diagram for a hydraulic jump (specific energy decreases).
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710 FLUID MECHANICS
3. The wall shear stress and the associated losses are negligible relative to the losses that occur during the hydraulic jump due to the intense agitation. 4. The channel is wide and horizontal. 5. There are no external or body forces other than gravity. For a channel of width b, the conservation of mass or continuity relation . . m2 ! m1 can be expressed as ry1bV1 ! ry2bV2 or y 1V1 ! y 2V2
(13–67)
Noting that the only forces acting on the control volume in the horizontal x→ #→ direction are the pressure forces, the momentum equation a F ! a bm V % out # → bm V in the x-direction becomes a balance between hydrostatic pressure a in
forces and momentum transfer,
# # P1, avg A 1 % P2, avg A 2 ! mV2 % mV1
(13–68)
where P1, avg ! rgy1/2 and P2, avg ! rgy2/2. For a channel width of b, we have . . . A1 ! y1b, A2 ! y2b, and m ! m 2 ! m 1 ! rA1V1 ! ry1bV1. Substituting and simplifying, the momentum equation reduces to y 21 % y 22 !
2y 1V1 (V2 % V1) g
(13–69)
Eliminating V2 by using V2 ! (y1/y2)V1 from the continuity equation gives y 21 % y 22 !
2y 1V 21 (y 1 % y 2) gy 2
(13–70)
Canceling the common factor y1 % y2 from both sides and rearranging give y2 2 y2 a b & % 2Fr 21 ! 0 y1 y1
(13–71)
where Fr1 ! V1, 1gy 1. This is a quadratic equation for y2/y1, and it has two roots—one negative and one positive. Noting that y2/y1 cannot be negative since both y2 and y1 are positive quantities, the depth ratio y2/y1 is determined to be Depth ratio:
y2 ! 0.5a%1 & 21 & 8Fr 21 b y1
(13–72)
The energy equation for this horizontal flow section can be expressed as y1 &
V 21 V 22 ! y2 & & hL 2g 2g
(13–73)
Noting that V2 ! (y1/y2)V1 and Fr1 ! V1 , 1gy 1, the head loss associated with hydraulic jump is expressed as hL ! y 1 % y 2 &
y 21 y 1Fr 21 V 21 % V 22 ! y1 % y2 & a1 % 2b 2g 2 y2
(13–74)
The energy line for a hydraulic jump is shown in Fig. 13–33. The drop in the energy line across the jump represents the head loss hL associated with the jump.
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711 CHAPTER 13
For a given Fr1 and y1, the downstream flow depth y2 and the head loss hL can be calculated from Eqs. 13–72 and 13–74, respectively. Plotting hL against Fr1 would reveal that hL becomes negative for Fr1 ( 1, which is impossible (it would correspond to negative entropy generation, which would be a violation of the second law of thermodynamics). Thus we conclude that the upstream flow must be supercritical (Fr1 ' 1) for a hydraulic jump to occur. In other words, it is impossible for subcritical flow to undergo a hydraulic jump. This is analogous to gas flow having to be supersonic (Mach number greater than 1) to undergo a shock wave. Head loss is a measure of the mechanical energy dissipated via internal fluid friction, and head loss is usually undesirable as it represents the mechanical energy wasted. But sometimes hydraulic jumps are designed in conjunction with stilling basins and spillways of dams, and it is desirable to waste as much of the mechanical energy as possible to minimize the mechanical energy of the water and thus its potential to cause damage. This is done by first producing supercritical flow by converting high pressure to high linear velocity, and then allowing the flow to agitate and dissipate part of its kinetic energy as it breaks down and decelerates to a subcritical velocity. Therefore, a measure of performance of a hydraulic jump is its fraction of energy dissipation. The specific energy of the liquid before the hydraulic jump is Es1 ! y1 & V 21/2g. Then the energy dissipation ratio (Fig. 13–34) can be expressed as Dissipation ratio !
hL hL hL ! ! E s1 y 1 & V 21,2g y 1(1 & Fr 21,2)
Energy line hL V 22 2g
V 12 2g V2 y1
y2
V1
(13–75)
The fraction of energy dissipation ranges from just a few percent for weak hydraulic jumps (Fr1 ( 2) to 85 percent for strong jumps (Fr1 ' 9). Unlike a normal shock in gas flow, which occurs at a cross section and thus has negligible thickness, the hydraulic jump occurs over a considerable channel length. In the Froude number range of practical interest, the length of the hydraulic jump is observed to be 4 to 7 times the downstream flow depth y2. Experimental studies indicate that hydraulic jumps can be considered in five categories as shown in Table 13–4, depending primarily on the value of the upstream Froude number Fr1. For Fr1 somewhat higher than 1, the liquid rises slightly during hydraulic jump, producing standing waves. At larger Fr1, highly damaging oscillating waves occur. The desirable range of Froude numbers is 4.5 ( Fr1 ( 9, which produces stable and well-balanced steady waves with high levels of energy dissipation within the jump. Hydraulic jumps with Fr1 ' 9 produce very rough waves. The depth ratio y2/y1 ranges from slightly over 1 for undular jumps that are mild and involve small rises in surface level to over 12 for strong jumps that are rough and involve high rises in surface level. In this section we limit our consideration to wide horizontal rectangular channels so that edge and gravity effects are negligible. Hydraulic jumps in nonrectangular and sloped channels behave similarly, but the flow characteristics and thus the relations for depth ratio, head loss, jump length, and dissipation ratio are different.
(1) Dissipation ratio !
(2) hL hL ! Es1 y1 & V12/2g
FIGURE 13–34 The energy dissipation ratio represents the fraction of mechanical energy dissipated during a hydraulic jump.
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712 FLUID MECHANICS
TA B L E 1 3 – 4 Classification of hydraulic jumps Source: U.S. Bureau of Reclamation (1955).
Upstream Fr1
Depth Ratio y2/y1
Fraction of Energy Dissipation
(1
1
0
1–2
(5%
1–1.7
Surface Profile
Description Impossible jump. Would violate the second law of thermodynamics. Undular jump (or standing wave). Small rise in surface level. Low energy dissipation. Surface rollers develop near Fr ! 1.7.
1.7–2.5
2–3.1
5–15%
2.5–4.5
3.1–5.9
15–45%
Oscillating jump. Pulsations caused by entering jets at the bottom generate large waves that can travel for miles and damage earth banks. Should be avoided in the design of stilling basins.
4.5–9
5.9–12
45–70%
Steady jump. Stable, well-balanced, and insensitive to downstream conditions. Intense eddy motion and high level of energy dissipation within the jump. Recommended range for design.
'9
'12
70–85%
Strong jump. Rough and intermittent. Very effective energy dissipation, but may be uneconomical compared to other designs.
Energy line
V1 ! 7 m/s
y2
y 1 ! 0.8 m (1)
(2)
FIGURE 13–35 Schematic for Example 13–7.
V2
V1
V2
y2
Weak jump. Surface rising smoothly, with small rollers. Low energy dissipation.
EXAMPLE 13–7
hL
y1
Hydraulic Jump
Water discharging into a 10-m-wide rectangular horizontal channel from a sluice gate is observed to have undergone a hydraulic jump. The flow depth and velocity before the jump are 0.8 m and 7 m/s, respectively. Determine (a) the flow depth and the Froude number after the jump, (b) the head loss and the dissipation ratio, and (c) the wasted power production potential due to the hydraulic jump (Fig. 13–35).
SOLUTION Water at a specified depth and velocity undergoes a hydraulic jump in a horizontal channel. The depth and Froude number after the jump,
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713 CHAPTER 13
the head loss and the dissipation ratio, and the wasted power potential are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Properties The density of water is 1000 kg/m3. Analysis (a) The Froude number before the hydraulic jump is
Fr1 !
V1 2gy 1
!
7 m/s 2(9.81 m/s2)(0.8 m)
! 2.50
which is greater than 1. Therefore, the flow is indeed supercritical before the jump. The flow depth, velocity, and Froude number after the jump are
y 2 ! 0.5y 1a%1 & 21 & 8Fr 21 b ! 0.5(0.8 m)a%1 & 21 & 8 $ 2.50 2 b ! 2.46 m V2 ! Fr2 !
y1 0.8 m V ! (7 m/s) ! 2.28 m/s y 2 1 2.46 m V2 2gy 2
!
2.28 m/s 2(9.81 m/s2)(2.46 m)
! 0.464
Note that the flow depth triples and the Froude number reduces to about one-fifth after the jump. (b) The head loss is determined from the energy equation to be
hL ! y 1 % y 2 &
V 21 % V 22 (7 m/s)2 % (2.28 m/s)2 ! (0.8 m) % (2.46 m) & 2g 2(9.81 m/s2)
! 0.572 m The specific energy of water before the jump and the dissipation ratio is
E s1 ! y 1 &
V 21 (7 m/s)2 ! (0.8 m) & ! 3.30 m 2g 2(9.81 m/s2)
Dissipation ratio !
hL 0.572 m ! ! 0.173 E s1 3.30 m
Therefore, 17.3 percent of the available head (or mechanical energy) of the liquid is wasted (converted to thermal energy) as a result of frictional effects during this hydraulic jump. (c) The mass flow rate of water is
# m# ! rV ! rby 1V1 ! (1000 kg/m3)(0.8 m)(10 m)(7 m/s) ! 56,000 kg/s
Then the power dissipation corresponding to a head loss of 0.572 m becomes
# E dissipated ! m# ghL ! (56,000 kg/s)(9.81 m/s2)(0.572 m)a ! 314,000 N . m/s ! 314 kW
1N b 1 kg . m/s2
Discussion The results show that the hydraulic jump is a highly dissipative process, wasting 314 kW of power production potential in this case. That is, if the water is routed to a hydraulic turbine instead of being released from
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714 FLUID MECHANICS
the sluice gate, up to 314 kW of power could be generated. But this potential is converted to useless thermal energy instead of useful power, causing a temperature rise of
# E dissipated 314 kJ/s ! 0.0013-C /T ! # ! mc (56,000 kg/s)(4.18 kJ/kg . -C) p
for water. Note that a 314-kW resistance heater would cause the same temperature rise for water flowing at a rate of 56,000 kg/s.
13–9
■
FLOW CONTROL AND MEASUREMENT
The flow rate in pipes and ducts is controlled by various kinds of valves. Liquid flow in open channels, however, is not confined, and thus the flow rate is controlled by partially blocking the channel. This is done by either allowing the liquid to flow over the obstruction or under it. An obstruction that allows the liquid to flow over it is called a weir, and an obstruction with an adjustable opening at the bottom that allows the liquid to flow underneath it is called an underflow gate. Such devices can be used to control the flow rate through the channel as well as to measure it.
Underflow Gates
There are numerous types of underflow gates to control the flow rate, each with certain advantages and disadvantages. Underflow gates are located at the bottom of a wall, dam, or an open channel. Two common types of such gates, the sluice gate and the drum gate, are shown in Fig. 13–36. A sluice gate is typically vertical and has a plane surface, whereas a drum gate has a circular cross section with a streamlined surface. When the gate is opened, the upstream liquid accelerates as it approaches the gate, reaches the critical speed at the gate, and accelerates further to supercritical speeds past the gate. Therefore, an underflow gate is analogous to a converging–diverging nozzle in gas dynamics. The discharge from an underflow gate is called a free outflow if the liquid jet streaming out of the gate is open to the atmosphere, and it is called a drowned (or submerged) outflow if the discharged liquid flashes back and submerges the jet, as shown in Fig. 13–36. In drowned flow, the liquid jet undergoes a hydraulic jump, and thus the downstream flow is subcritical. Also, drowned outflow involves a high level of turbulence and backflow, and thus a large head loss hL. The flow depth-specific energy diagram for flow through underflow gates with free and drowned outflow is given in Fig. 13–37. Note that the specific energy remains constant for idealized gates with negligible frictional effects (from point 1 to point 2a), but decreases for actual gates. The downstream is supercritical for a gate with free outflow (point 2b), but subcritical for one with drowned outflow (point 2c) since a drowned outflow also involves a hydraulic jump to subcritical flow, which involves considerable mixing and energy dissipation. Assuming the frictional effects to be negligible and the upstream (or reservoir) velocity to be low, it can be shown by using the Bernoulli equation that the discharge velocity of a free jet is (see Chap. 5 for details) V ! 22gy 1
(13–76)
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715 CHAPTER 13 Sluice gate
Sluice gate
Drum V1
y1
y1
Vena contracta a
V1
V1 y2
V2
y2
y1
V2
y2
a
(a) Sluice gate with free outflow
V2
(c) Drum gate
(b) Sluice gate with drowned outflow
FIGURE 13–36 Common types of underflow gates to control flow rate.
The frictional effects can be accounted for by modifying this relation with a discharge coefficient Cd. Then the discharge velocity at the gate and the flow rate become V ! C d 22gy 1
# V ! C d ba22gy 1
and
(13–77)
where b and a are the width and the height of the gate opening, respectively. The discharge coefficient Cd ! 1 for idealized flow, but Cd ( 1 for actual flow through the gates. Experimentally determined values of Cd for underflow gates are plotted in Fig. 13–38 as functions of the contraction coefficient y2/a and the depth ratio y1/a. Note that most values of Cd for free outflow from a vertical sluice gate range between 0.5 and 0.6. The Cd values drop sharply for drowned outflow, as expected, and the flow rate decreases for the same upstream conditions. For a given value of y1/a, the value of Cd decreases with increasing y2/a. 0.6 Free outflow
Subcritical flow
y
Es1 ! y1 &
V 21 2g
1 Frictionless gate
2c Drowned outflow
Supercritical flow 2b 2a
Es
Es1 ! Es2a
FIGURE 13–37 Schematic and flow depth-specific energy diagram for flow through underflow gates.
0.5
0.4
Cd 0.3
0.2 Drowned outflow
FIGURE 13–38 Discharge coefficients for drowned and free discharge from underflow gates.
0.1 y 2 /a ! 2 0
0
2
3
4 4
5
6 6
7
8 8 y 1/a
10
12
14
16
From Henderson, Open Channel Flow, 1st Edition, © 1966. Reprinted by permission of Pearson Education, Inc., Upper Saddle River, NJ.
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716 FLUID MECHANICS Sluice gate
EXAMPLE 13–8
Sluice Gate with Drowned Outflow
Water is released from a 3-m-deep reservoir into a 6-m-wide open channel through a sluice gate with a 0.25-m-high opening at the channel bottom. The flow depth after all turbulence subsides is measured to be 1.5 m. Determine the rate of discharge (Fig. 13–39).
y1 ! 3 m
SOLUTION Water is released from a reservoir through a sluice gate into an y 2 ! 1.5 m a ! 0.25 m
FIGURE 13–39 Schematic for Example 13–8.
open channel. For specified flow depths, the rate of discharge is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Analysis The depth ratio y1/a and the contraction coefficient y2/a are
y1 3m ! ! 12 a 0.25 m
and
y2 1.5 m ! !6 a 0.25 m
The corresponding discharge coefficient is determined from Fig. 13–38 to be Cd ! 0.47. Then the discharge rate becomes
# V ! C d ba22gy 1 ! 0.47(6 m)(0.25 m)22(9.81 m/s2)(3 m) ! 5.41 m3/s
Discussion In the case of free flow, the discharge coefficient would be Cd ! 0.59, with a corresponding flow rate of 6.78 m3/s. Therefore, the flow rate decreases considerably when the outflow is drowned.
Overflow Gates
Recall that the total mechanical energy of a liquid at any cross section of an open channel can be expressed in terms of heads as H ! zb & y & V 2/2g, where y is the flow depth, zb is the elevation of the channel bottom, and V is the average flow velocity. During flow with negligible frictional effects (head loss hL ! 0), the total mechanical energy remains constant, and the one-dimensional energy equation for open-channel flow between upstream section 1 and downstream section 2 can be written as z b1 & y 1 &
y . V ! constant
Subcritical flow, Fr ( 1
V2 2g y yc
Fr ! 1
Es ! y
Critical depth
Supercritical flow, Fr ' 1
Emin
FIGURE 13–40 Variation of specific energy Es with depth y for a specified flow rate.
V 21 V 22 ! z b2 & y 2 & 2g 2g
E s1 ! /z b & E s2
(13–78)
where Es ! y & V 2/2g is the specific energy and /zb ! zb2 % zb1 is the elevation of the bottom point of flow at section 2 relative to that at section 1. Therefore, the specific energy of a liquid stream increases by |/zb| during downhill flow (note that /zb is negative for channels inclined down), decreases by /zb during uphill flow, and remains constant during horizontal flow. (The specific energy also decreases by hL for all cases if the frictional effects are not negligible.) . For a channel. of constant width b, V ! AcV ! byV ! constant in steady flow and V ! V /Ac. Then the specific energy can be expressed as Es ! y &
Es
or
# V2 2gb 2y 2
(13–79)
The variation of the specific energy Es with flow depth y for steady flow in a channel of constant width b is replotted in Fig. 13–40. This diagram is extremely valuable as it shows the allowable states during flow. Once the upstream conditions at a flow section 1 are specified, the state of the liquid
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717 CHAPTER 13
at any section 2 on an Es-y diagram must fall on a point on the specific energy curve that passes through point 1.
Flow over a Bump with Negligible Friction
Now consider steady flow with negligible friction over a bump of height /zb in a horizontal channel of constant width b, as shown in Fig. 13–41. The energy equation in this case is, from Eq. 13–78, E s2 ! E s1 % /z b
(13–80)
Therefore, the specific energy of the liquid decreases by /zb as it flows over the bump, and the state of the liquid on the Es-y diagram shifts to the left by /zb, as shown in Fig. 13–41. The continuity equation for a channel of large width is y2V2 ! y1V1 and thus V2 ! (y1/y2)V1. Then the specific energy of the liquid over the bump can be expressed as E s2 ! y 2 &
V 22 2g
→
E s1 % /z b ! y 2 &
V 21 y 21 2g y 22
(13–81)
Rearranging, y 32 % (E s1 % /z b)y 22 &
V 21 2 y1 ! 0 2g
(13–82)
which is a third-degree polynomial equation in y2 and thus has three solutions. Disregarding the negative solution, it appears that the flow depth over the bump can have two values. Now the curious question is, does the liquid level rise or drop over the bump? Our intuition says the entire liquid body will follow the bump and thus the liquid surface will rise over the bump, but this is not necessarily so. Noting that specific energy is the sum of the flow depth and dynamic head, either scenario is possible, depending on how the velocity changes. The Es-y diagram in Fig. 13–41 gives us the definite answer: If the flow before the bump is subcritical (state 1a), the flow depth y2 decreases (state 2a). If the decrease in flow depth is greater than the bump height (i.e., y1 % y2 ' /zb), the free surface is suppressed. But if the flow is supercritical as it approaches the bump (state 1b), the flow depth rises over the bump (state 2b), creating a bigger bump over the free surface.
y
1a Subcritical flow
2a
/zb 2b Emin ! Ec
Supercritical flow 1b Es
Supercritical upstream flow
y1
y2
V1
Subcritical upstream flow V2
/zb Bump
FIGURE 13–41 Schematic and flow depth-specific energy diagram for flow over a bump for subcritical and supercritical upstream flows.
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718 FLUID MECHANICS
The situation is reversed if the channel has a depression of depth /zb instead of a bump: The specific energy in this case increases (so that state 2 is to the right of state 1 on the Es-y diagram) since /zb is negative. Therefore, the flow depth increases if the approach flow is subcritical and decreases if it is supercritical. Now let’s reconsider flow over a bump with negligible friction, as discussed earlier. As the height of the bump /zb is increased, point 2 (either 2a or 2b for sub- or supercritical flow) continues shifting to the left on the Es-y diagram, until finally reaching the critical point. That is, the flow over the bump is critical when the bump height is /zc ! Es1 % Esc ! Es1 % Emin, and the specific energy of the liquid reaches its minimum level. The question that comes to mind is, what happens if the bump height is increased further? Does the specific energy of the liquid continue decreasing? The answer to this question is a resounding no since the liquid is already at its minimum energy level, and its energy cannot decrease any further. In other words, the liquid is already at the furthest left point on the Esy diagram, and no point further left can satisfy conservation of mass, momentum, and energy. Therefore, the flow must remain critical. The flow at this state is said to be choked. In gas dynamics, this is analogous to the flow in a converging nozzle accelerating as the back pressure is lowered, and reaching the speed of sound at the nozzle exit when the back pressure reaches the critical pressure. But the nozzle exit velocity remains at the sonic level no matter how much the back pressure is lowered. Here again, the flow is choked.
Broad-Crested Weir
The discussions on flow over a high bump can be summarized as follows: The flow over a sufficiently high obstruction in an open channel is always critical. Such obstructions placed intentionally in an open channel to measure the flow rate are called weirs. Therefore, the flow velocity over a sufficiently broad weir is the critical velocity, which is expressed as V ! 1gy c, where yc is the critical depth. Then the flow rate over a weir of width b can be expressed as # V ! A cV ! y cb2gy c ! bg 1,2y 3,2 c
Discharge V1
H
Pw
Vc
yc
Broad-crested weir Lw
FIGURE 13–42 Flow over a broad-crested weir.
(13–83)
A broad-crested weir is a rectangular block of height Pw and length Lw that has a horizontal crest over which critical flow occurs (Fig. 13–42). The upstream head above the top surface of the weir is called the weir head and is denoted by H. To obtain a relation for the critical depth yc in terms of weir head H, we write the energy equation between a section upstream and a section over the weir for flow with negligible friction as H & Pw &
V 2c V 21 ! y c & Pw & 2g 2g
(13–84)
Cancelling Pw from both sides and substituting Vc ! 1gy c give yc !
V 21 2 aH & b 3 2g
(13–85)
Substituting into Eq. 13–83, the flow rate for this idealized flow case with negligible friction is determined to be
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719 CHAPTER 13
# V 21 3,2 2 3,2 V ideal ! b2ga b aH & b 3 2g
(13–86)
This relation shows the functional dependence of the flow rate on the flow parameters, but it overpredicts the flow rate by several percent because it does not consider the frictional effects. These effects are properly accounted for by modifying the theoretical relation (Eq. 13–86) by an experimentally determined weir discharge coefficient Cwd as Broad-crested weir:
# V 21 3,2 2 3,2 V ! C wd, broadb2g a b aH & b 3 2g
(13–87)
where reasonably accurate values of discharge coefficients for broad-crested weirs can be obtained from (Chow, 1959) C wd, broad !
0.65 21 & H,Pw
(13–88)
More accurate but complicated relations for Cwd, broad are also available in the literature (e.g., Ackers, 1978). Also, the upstream velocity V1 is usually very low, and it can be disregarded. This is especially the case for high weirs. Then the flow rate can be approximated as Broad-crested weir with low V1:
# 2 3,2 V ! C wd, broadb2ga b H 3,2 3
(13–89)
It should always be kept in mind that the basic requirement for the use of Eqs. 13–87 to 13–89 is the establishment of critical flow above the weir, and this puts some limitations on the weir length Lw. If the weir is too long (Lw ' 12H), wall shear effects dominate and cause the flow over the weir to be subcritical. If the weir is too short (Lw ( 2H), the liquid may not be able to accelerate to critical velocity. Based on observations, the proper length of the broad-crested weir is 2H ( Lw ( 12H. Note that a weir that is too long for one flow may be too short for another flow, depending on the value of the weir head H. Therefore, the range of flow rates should be known before a weir can be selected.
Sharp-Crested Weirs
A sharp-crested weir is a vertical plate placed in a channel that forces the liquid to flow through an opening to measure the flow rate. The type of the weir is characterized by the shape of the opening. A vertical thin plate with a straight top edge is referred to as rectangular weir since the cross section of the flow over it is rectangular; a weir with a triangular opening is referred to as a triangular weir; etc. Upstream flow is subcritical and becomes critical as it approaches the weir. The liquid continues to accelerate and discharges as a supercritical flow stream that resembles a free jet. The reason for acceleration is the steady decline in the elevation of the free surface, and the conversion of this elevation head into velocity head. The flow-rate correlations given below are based on the free overfall of liquid discharge past the weir, called nappes, being clear from the weir. It may be necessary to ventilate the space under the nappe to assure atmospheric pressure underneath. Empirical relations for drowned weirs are also available.
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720 FLUID MECHANICS
H V1
u2(h)
h
Nappe
2 1 Pw
y x
(1)
Weir (2)
FIGURE 13–43 Flow over a sharp-crested weir.
Consider the flow of a liquid over a sharp-crested weir placed in a horizontal channel, as shown in Fig. 13–43. For simplicity, the velocity upstream of the weir is approximated as being nearly constant through vertical cross section 1. The total energy of the upstream liquid expressed as a head relative to the channel bottom is the specific energy, which is the sum of the flow depth and the velocity head. That is, y1 & V 12/2g, where y1 ! H & Pw. The flow over the weir is not one-dimensional since the liquid undergoes large changes in velocity and direction over the weir. But the pressure within the nappe is atmospheric. A simple relation for the variation of liquid velocity over the weir can be obtained by assuming negligible friction and writing the Bernoulli equation between a point in upstream flow (point 1) and a point over the weir at a distance h from the upstream liquid level as H & Pw &
V 21 u 22 ! (H & Pw % h) & 2g 2g
(13–90)
Cancelling the common terms and solving for u2, the idealized velocity distribution over the weir is determined to be u 2 ! 22gh & V 21
(13–91)
In reality, water surface level drops somewhat over the weir as the water starts its free overfall (the drawdown effect at the top) and the flow separation at the top edge of the weir further narrows the nappe (the contraction effect at the bottom). As a result, the flow height over the weir is considerably smaller than H. When the drawdown and contraction effects are disregarded for simplicity, the flow rate is obtained by integrating the product of the flow velocity and the differential flow area over the entire flow area, # V!
$
Ac
u 2 dA c2 !
$
H
22gh & V 21 w dh
(13–92)
h!0
where w is the width of the flow area at distance h from the upstream free surface. In general, w is a function of h. But for a rectangular weir, w ! b, which is constant. Then the integration can be performed easily, and the flow rate for a rectangular weir for idealized flow with negligible friction and negligible drawdown and contraction effects is determined to be # V 21 3,2 V 21 3,2 2 V ideal ! b22gcaH & b % a b d 3 2g 2g
(13–93)
When the weir height is large relative to the weir head (Pw '' H), the upstream velocity V1 is low and the upstream velocity head can be neglected. That is, V 12/2g (( H. Then, # 2 V ideal, rec ! b22gH 3,2 3
(13–94)
Therefore, the flow rate can be determined from a knowledge of two geometric quantities: the crest width b and the weir head H, which is the vertical distance between the weir crest and the upstream free surface. This simplified analysis gives the general form of the flow-rate relation, but it needs to be modified to account for the frictional and surface tension effects, which play a secondary role as well as the drawdown and contraction
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721 CHAPTER 13
effects. Again this is done by multiplying the ideal flow-rate relations by an experimentally determined weir discharge coefficient Cwd. Then the flow rate for a sharp-crested rectangular weir is expressed as Sharp-crested rectangular weir:
# 2 V rec ! C wd, rec b22gH 3,2 3
Upstream free surface
(13–95)
where, from Ref. 1 (Ackers, 1978), H C wd, rec ! 0.598 & 0.0897 Pw
for
H 02 Pw
(13–96)
This formula is applicable over a wide range of upstream Reynolds numbers defined as Re ! V1H/n. More precise but also more complex correlations are also available in the literature. Note that Eq. 13–95 is valid for full-width rectangular weirs. If the width of the weir is less than the channel width so that the flow is forced to contract, an additional coefficient for contraction correction should be incorporated to properly account for this effect. Another type of sharp-crested weir commonly used for flow measurement is the triangular weir (also called the V-notch weir) shown in Fig. 13–44. The triangular weir has the advantage that it maintains a high weir head H even for small flow rates because of the decreasing flow area with decreasing H, and thus it can be used to measure a wide range of flow rates accurately. From geometric consideration, the notch width can be expressed as w ! 2(H % h) tan(u/2), where u is the V-notch angle. Substituting into Eq. 13–92 and performing the integration give the ideal flow rate for a triangular weir to be # 8 u V ideal, tri ! tana b 22gH 5,2 15 2
H
# 8 u V ! C wd, tri tana b 22gH 5,2 15 2
u Weir plate
Pw
FIGURE 13–44 A triangular (or V-notch) sharp-crested weir plate geometry. The view is from downstream looking upstream.
Depression over the bump
(13–97)
(13–98)
where the values of Cwd typically range between 0.58 and 0.62. Therefore, the fluid friction, the constriction of flow area, and other dissipative effects cause the flow rate through the V-notch to decrease by about 40 percent compared to the ideal case. For most practical cases (H ' 0.2 m and 45° ( u ( 120°), the value of the weir discharge coefficient is about Cwd ! 0.58. More precise values are available in the literature.
y2
y1 ! 0.80 m
where we again neglected the upstream velocity head. The frictional and other dissipative effects are accounted for conveniently by multiplying the ideal flow rate by a weir discharge coefficient. Then the flow rate for a sharp-crested triangular weir becomes Sharp-crested triangular weir:
h
w
/zb ! 0.15 m
Bump
V1 ! 1.2 m/s y
1
y1 y2
Subcritical flow
2
/zb
EXAMPLE 13–9
Subcritical Flow over a Bump
Water flowing in a wide horizontal open channel encounters a 15-cm-high bump at the bottom of the channel. If the flow depth is 0.80 m and the velocity is 1.2 m/s before the bump, determine if the water surface is depressed over the bump (Fig. 13–45).
SOLUTION Water flowing in a horizontal open channel encounters a bump. It will be determined if the water surface is depressed over the bump.
Es2
Es1
Es
FIGURE 13–45 Schematic and flow depth-specific energy diagram for Example 13–9.
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722 FLUID MECHANICS
Assumptions 1 The flow is steady. 2 Frictional effects are negligible so that there is no dissipation of mechanical energy. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis The upstream Froude number and the critical depth are
Fr1 !
V1 2gy 1
!
1.2 m/s 2(9.81 m2/s)(0.80 m)
! 0.428
# (by 1V1)2 1,3 y 21V 21 1,3 (0.8 m)2(1.2 m,s)2 1,3 V 2 1,3 y c ! a 2b ! a b b ! a ! a b ! 0.455 m g gb gb 2 9.81 m/s2 The flow is subcritical since Fr ( 1 and therefore the flow depth decreases over the bump. The upstream specific energy is
E s1 ! y 1 &
V 21 (1.2 m/s)2 ! (0.80 m) & ! 0.873 m 2g 2(9.81 m/s2)
The flow depth over the bump can be determined from
y 32 % (E s1 % /z b)y 22 &
V 21 2 y1 ! 0 2g
Substituting,
y 32 % (0.873 % 0.15 m)y 22 &
(1.2 m/s)2 (0.80 m)2 ! 0 2(9.81 m/s2)
or
y 32 % 0.723y 22 & 0.0470 ! 0 Using an equation solver, the three roots of this equation are determined to be 0.59 m, 0.36 m, and %0.22 m. We discard the negative solution as physically impossible. We also eliminate the solution 0.36 m since it is less than the critical depth, and it can occur only in supercritical flow. Thus the only meaningful solution for flow depth over the bump is y2 ! 0.59 m. Then the distance of the water surface over the bump from the channel bottom is /zb & y2 ! 0.15 & 0.59 ! 0.74 m, which is less than y1 ! 0.80 m. Therefore, the water surface is depressed over the bump in the amount of
Depression ! y 1 % (y 2 & /z b) ! 0.80 % (0.59 & 0.15) ! 0.06 m
b! 5 m
Discussion Note that having y2 ( y1 does not necessarily indicate that the water surface is depressed (it may still rise over the bump). The surface is depressed over the bump only when the difference y1 % y2 is larger than the bump height /zb. Also, the actual value of depression may be different than 0.06 m because of the frictional effects that are neglected in the analysis.
y1 ! 1.5 m V1
EXAMPLE 13–10 Pw ! 0.60 m
Sharp-crested rectangular weir
FIGURE 13–46 Schematic for Example 13–10.
Measuring Flow Rate by a Weir
The flow rate of water in a 5-m-wide horizontal open channel is being measured with a 0.60-m-high sharp-crested rectangular weir of equal width. If the water depth upstream is 1.5 m, determine the flow rate of water (Fig. 13–46).
SOLUTION The water depth upstream of a horizontal open channel equipped with a sharp-crested rectangular weir is measured. The flow rate is to be determined.
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723 CHAPTER 13
Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis The weir head is
H ! y 1 % Pw ! 1.5 % 0.60 ! 0.90 m The discharge coefficient of the weir is
C wd, rec ! 0.598 & 0.0897
H 0.90 ! 0.733 ! 0.598 & 0.0897 Pw 0.60
The condition H/Pw ( 2 is satisfied since 0.9/0.6 ! 1.5. Then the water flow rate through the channel becomes
# 2 V rec ! C wd, rec b22gH 3,2 3 2 ! (0.733) (5 m)22(9.81 m/s2)(0.90 m)3,2 3 ! 9.24 m3/s Discussion The upstream velocity and the upstream velocity head are
# V 9.24 m3/s V1 ! ! ! 1.23 m/s by 1 (5 m)(1.5 m)
and
V 21 (1.23 m/s)2 ! ! 0.077 m 2g 2(9.81 m/s2)
This is 8.6 percent of the weir head, which is significant. When the upstream velocity head is considered, the flow rate becomes 10.2 m3/s, which is about 10 percent higher than the value determined. Therefore, it is good practice to consider the upstream velocity head unless the weir height Pw is very large relative to the weir head H.
SUMMARY Open-channel flow refers to the flow of liquids in channels open to the atmosphere or in partially filled conduits. The flow in a channel is said to be uniform if the flow depth (and thus the average velocity) remains constant. Otherwise, the flow is said to be nonuniform or varied. The hydraulic radius is defined as Rh ! Ac/p. The dimensionless Froude number is defined as Fr !
V
!
V
2gL c 2gy The flow is classified as subcritical for Fr ( 1, critical for Fr ! 1, and supercritical for Fr ' 1. Flow depth in critical flow is called the critical depth and is expressed as # # V2 V 2 1,3 yc ! 2 or y c ! a 2b gA c gb
where b is the channel width for wide channels. The speed at which a surface disturbance travels through a liquid of depth y is the wave speed c0, which is expressed as
c0 ! 1gy. The total mechanical energy of a liquid in a channel is expressed in terms of heads as H ! zb & y &
V2 2g
where zb is the elevation head, P/rg ! y is the pressure head, and V 2/2g is the velocity head. The sum of the pressure and dynamic heads is called the specific energy Es, Es ! y &
V2 2g
The continuity equation is Ac1V1 ! Ac2V2. The energy equation is expressed as y1 &
V 21 V 22 & S 0L ! y 2 & & hL 2g 2g
Here hL is the head loss and S0 ! tan u is the bottom slope of a channel. The friction slope is defined as Sf ! hL/L.
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724 FLUID MECHANICS
The flow depth in uniform flow is called the normal depth yn, and the average flow velocity is called the uniform-flow velocity V0. The velocity and flow rate in uniform flow are given by a V0 ! R 2,3 S 1,2 n h 0
and
# a 1,2 V ! A cR 2,3 h S0 n
where n is the Manning coefficient whose value depends on the roughness of the channel surfaces, and a ! 1 m1/3/s ! (3.2808 ft)1/3/s ! 1.486 ft1/3/s. If yn ! yc, the flow is uniform critical flow, and the bottom slope S0 equals the critical slope Sc expressed as Sc !
gn 2y c a 2R 4,3 h
which simplifies to
Sc !
gn 2 a 2y 1,3 c
for film flow or flow in a wide rectangular channel with b '' yc. The best hydraulic cross section for an open channel is the one with the maximum hydraulic radius, or equivalently, the one with the minimum wetted perimeter for a specified cross section. The criteria for best hydraulic cross section for a rectangular channel is y ! b/2. The best cross section for trapezoidal channels is half of a hexagon. In rapidly varied flow (RVF), the flow depth changes markedly over a relatively short distance in the flow direction. Any change from supercritical to subcritical flow occurs through a hydraulic jump, which is a highly dissipative process. The depth ratio y2/y1, head loss, and energy dissipation ratio during hydraulic jump are expressed as y2 ! 0.5a%1 & 21 & 8Fr 21 b y1 hL ! y 1 % y 2 & ! y1 % y2 & Dissipation ratio ! !
V 21
% 2g
V 22
y 1Fr 21 y 21 a1 % 2b 2 y2
hL hL ! E s1 y 1 & V 21,2g hL y 1(1 & Fr 21 ,2)
An obstruction that allows the liquid to flow over it is called a weir, and an obstruction with an adjustable opening at the bottom that allows the liquid to flow underneath it is called an underflow gate. The flow rate through a sluice gate is given by # V ! C d ba22gy 1 where b and a are the width and the height of the gate opening, respectively, and Cd is the discharge coefficient, which accounts for the frictional effects. A broad-crested weir is a rectangular block that has a horizontal crest over which critical flow occurs. The upstream head above the top surface of the weir is called the weir head, H. The flow rate is expressed as # V 21 3,2 2 3,2 V ! C wd, broadb2ga b aH & b 3 2g
where the discharge coefficient is C wd, broad !
0.65
21 & H,Pw The flow rate for a sharp-crested rectangular weir is expressed as # 2 V rec ! C wd, rec b22gH 3,2 3 where H C wd, rec ! 0.598 & 0.0897 Pw
for
H 02 Pw
For a sharp-crested triangular weir, the flow rate is given as # 8 u V ! C wd, tri tana b 22gH 5,2 15 2
where the values of Cwd, tri typically range between 0.58 and 0.62. Open-channel analysis is commonly used in the design of sewer systems, irrigation systems, floodways, and dams. Some open-channel flows are analyzed in Chap. 15 using computational fluid dynamics (CFD).
REFERENCES AND SUGGESTED READING 1. P. Ackers et al. Weirs and Flumes for Flow Measurement. New York: Wiley, 1978.
4. V. T. Chow. Open Channel Hydraulics. New York: McGraw-Hill, 1959.
2. B. A. Bakhmeteff. Hydraulics of Open Channels. New York: McGraw-Hill, 1932.
5. C. T. Crowe, J. A. Roberson, and D. F. Elger. Engineering Fluid Mechanics, 7th ed. New York: Wiley, 2001.
3. M. H. Chaudhry. Open Channel Flow. Upper Saddle River, NJ: Prentice Hall, 1993.
6. R. H. French. Open Channel Hydraulics. New York: McGraw-Hill, 1985.
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7. F. M. Henderson. Open Channel Flow. New York: Macmillan, 1966.
11. F. M. White. Fluid Mechanics, 5th ed. New York: McGraw-Hill, 2003.
8. C. C. Mei. The Applied Dynamics of Ocean Surface Waves. New York: Wiley, 1983.
12. U. S. Bureau of Reclamation. “Research Studies on Stilling Basins, Energy Dissipaters, and Associated Appurtenances,” Hydraulic Lab Report Hyd.-399, June 1, 1955.
9. B. R. Munson, D. F. Young, and T. H. Okiishi. Fundamentals of Fluid Mechanics, 4th ed. New York: Wiley, 2002. 10. M. C. Potter and D. C. Wiggert. Mechanics of Fluids, 2nd ed. Upper Saddle River, NJ: Prentice Hall, 1997.
PROBLEMS* Classification, Froude Number, and Wave Speed 13–1C flow?
How does open-channel flow differ from internal
13–2C What is the driving force for flow in an open channel? How is the flow rate in an open channel established? 13–3C How does the pressure change along the free surface in an open-channel flow? 13–4C Consider steady fully developed flow in an open channel of rectangular cross section with a constant slope of 5° for the bottom surface. Will the slope of the free surface also be 5°? Explain. 13–5C How does uniform flow differ from nonuniform flow in open channels? In what kind of channels is uniform flow observed? 13–6C What is normal depth? Explain how it is established in open channels. 13–7C What causes the flow in an open channel to be varied (or nonuniform)? How does rapidly varied flow differ from gradually varied flow?
13–11C What is critical depth in open-channel flow? For a given average flow velocity, how is it determined? 13–12C The flow in an open channel is observed to have undergone a hydraulic jump. Is the flow upstream from the jump necessarily supercritical? Is the flow downstream from the jump necessarily subcritical? 13–13 Consider the flow of water in a wide channel. Determine the speed of a small disturbance in the flow if the flow depth is (a) 10 cm and (b) 80 cm. What would your answer be if the fluid were oil? 13–14 Water at 20°C is flowing uniformly in a wide rectangular channel at an average velocity of 2 m/s. If the water depth is 0.2 m, determine (a) whether the flow is laminar or turbulent and (b) whether the flow is subcritical or supercritical. 13–15 Water at 20°C flows in a partially full 2-m-diameter circular channel at an average velocity of 2 m/s. If the maximum water depth is 0.5 m, determine the hydraulic radius, the Reynolds number, and the flow regime.
13–8C In open channels, how is hydraulic radius defined? Knowing the hydraulic radius, how can the hydraulic diameter of the channel be determined?
R!1m
13–9C Given the average flow velocity and the flow depth, explain how you would determine if the flow in open channels is tranquil, critical, or rapid. 13–10C What is the Froude number? How is it defined? What is its physical significance? * Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
0.5 m
FIGURE P13–15 13–16 Water at 15°C is flowing uniformly in a 2-m-wide rectangular channel at an average velocity of 4 m/s. If the water depth is 8 cm, determine whether the flow is subcritical or supercritical. Answer: supercritical 13–17 After heavy rain, water flows on a concrete surface at an average velocity of 1.3 m/s. If the water depth is 2 cm, determine whether the flow is subcritical or supercritical.
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13–18E Water at 70°F is flowing uniformly in a wide rectangular channel at an average velocity of 6 ft/s. If the water depth is 0.5 ft, determine (a) whether the flow is laminar or turbulent and (b) whether the flow is subcritical or supercritical. 13–19 Water at 10°C flows in a 3-m-diameter circular channel half-full at an average velocity of 2.5 m/s. Determine the hydraulic radius, the Reynolds number, and the flow regime (laminar or turbulent). 13–20 A single wave is initiated in a sea by a strong jolt during an earthquake. Taking the average water depth to be 2 km and the density of seawater to be 1.030 kg/m3, determine the speed of propagation of this wave.
Specific Energy and the Energy Equation 13–21C How is the specific energy of a fluid flowing in an open channel defined in terms of heads? 13–22C Consider steady flow of water through two identical open rectangular channels at identical flow rates. If the flow in one channel is subcritical and in the other supercritical, can the specific energies of water in these two channels be identical? Explain. 13–23C For a given flow rate through an open channel, the variation of specific energy with flow depth is studied. One person claims that the specific energy of the fluid will be minimum when the flow is critical, but another person claims that the specific energy will be minimum when the flow is subcritical. What is your opinion? 13–24C Consider steady supercritical flow of water through an open rectangular channel at a constant flow rate. Someone claims that the larger is the flow depth, the larger the specific energy of water. Do you agree? Explain. 13–25C During steady and uniform flow through an open channel of rectangular cross section, a person claims that the specific energy of the fluid remains constant. A second person claims that the specific energy decreases along the flow because of the frictional effects and thus head loss. With which person do you agree? Explain. 13–26C How is the friction slope defined? Under what conditions is it equal to the bottom slope of an open channel? 13–27C Consider steady flow of a liquid through a wide rectangular channel. It is claimed that the energy line of flow is parallel to the channel bottom when the frictional losses are negligible. Do you agree? 13–28C Consider steady one-dimensional flow through a wide rectangular channel. Someone claims that the total mechanical energy of the fluid at the free surface of a cross section is equal to that of the fluid at the channel bottom of the same cross section. Do you agree? Explain. 13–29C How is the total mechanical energy of a fluid during steady one-dimensional flow through a wide rectangular
channel expressed in terms of heads? How is it related to the specific energy of the fluid? 13–30C Express the one-dimensional energy equation for open-channel flow between an upstream section 1 and downstream section 2, and explain how the head loss can be determined. 13–31 Water flows steadily in a 0.8-m-wide rectangular channel at a rate of 0.7 m3/s. If the flow depth is 0.25 m, determine the flow velocity and if the flow is subcritical or supercritical. Also determine the alternate flow depth if the character of flow were to change. 13–32 Water at 15°C flows at a depth of 0.4 m with an average velocity of 6 m/s in a rectangular channel. Determine the specific energy of water and whether the flow is subcritical or supercritical. 13–33
Water at 15°C flows at a depth of 0.4 m with an average velocity of 6 m/s in a rectangular channel. Determine (a) the critical depth, (b) the alternate depth, and (c) the minimum specific energy. 13–34 Water at 10°C flows in a 6-m-wide rectangular channel at a depth of 0.55 m and a flow rate of 12 m3/s. Determine (a) the critical depth, (b) whether the flow is subcritical or supercritical, and (c) the alternate depth. Answers: (a) 0.742 m, (b) supercritical, (c) 1.03 m
13–35E Water at 65°F flows at a depth of 0.8 ft with an average velocity of 14 ft/s in a wide rectangular channel. Determine (a) the Froude number, (b) the critical depth, and (c) whether the flow is subcritical or supercritical. What would your response be if the flow depth were 0.2 ft? 13–36E 10 ft/s.
Repeat Prob. 13–35E for an average velocity of
13–37 Water flows through a 4-m-wide rectangular channel with an average velocity of 5 m/s. If the flow is critical, determine the flow rate of water. Answer: 51.0 m3/s 13–38 Water flows half-full through a 50-cm-diameter steel channel at an average velocity of 2.8 m/s. Determine the volume flow rate and whether the flow is subcritical or supercritical. 13–39 Water flows half-full through a hexagon channel of bottom width 2 m at a rate of 45 m3/s. Determine (a) the average velocity and (b) whether the flow is subcritical and supercritical. 13–40 Repeat Prob. 13–39 for a flow rate of 30 m3/s.
Uniform Flow and Best Hydraulic Cross Sections 13–41C When is the flow in an open channel said to be uniform? Under what conditions will the flow in an open channel remain uniform? 13–42C Consider uniform flow through a wide rectangular channel. If the bottom slope is increased, the flow depth will (a) increase, (b) decrease, or (c) remain constant.
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13–43C During uniform flow in an open channel, someone claims that the head loss can be determined by simply multiplying the bottom slope by the channel length. Can it be this simple? Explain. 13–44C During uniform flow in open channels, the flow velocity and the flow rate can be determined from the. 1/2 Manning equations expressed as V0 ! (a/n)R2/3 h S 0 and V 1/2. What is the value and dimension of the conS ! (a/n)AcR2/3 h 0 stant a in these equations in SI units? Also, explain how the Manning coefficient n can be determined when the friction factor f is known. 13–45C Show that for uniform critical flow, the general gn 2y c gn 2 critical slope relation S c ! 2 4,3 reduces to S c ! 2 1,3 for a Rh a yc film flow with b '' yc. 13–46C Which is a better hydraulic cross section for an open channel: one with a small or a large hydraulic radius? 13–47C Which is the best hydraulic cross section for an open channel: (a) circular, (b) rectangular, (c) trapezoidal, or (d) triangular? 13–48C The best hydraulic cross section for a rectangular open channel is one whose fluid height is (a) half, (b) twice, (c) equal to, or (d) one-third the channel width. 13–49C The best hydraulic cross section for a trapezoidal channel of base width b is one for which the length of the side edge of the flow section is (a) b, (b) b/2, (c) 2b, or (d) 13b. 13–50C Consider uniform flow through an open channel lined with bricks with a Manning coefficient of n ! 0.015. If the Manning coefficient doubles (n ! 0.030) as a result of some algae growth on surfaces while the flow cross section remains constant, the flow rate will (a) double, (b) decrease by a factor of 12, (c) remain unchanged, (d) decrease by half, or (e) decrease by a factor of 21/3. 13–51 Water is flowing uniformly in a finished-concrete channel of trapezoidal cross section with a bottom width of 0.6 m, trapezoid angle of 50°, and a bottom angle of 0.4°. If the flow depth is measured to be 0.45 m, determine the flow rate of water through the channel.
13–53E A 6-ft-diameter semicircular channel made of unfinished concrete is to transport water to a distance of 1 mi uniformly. If the flow rate is to reach 150 ft3/s when the channel is full, determine the minimum elevation difference across the channel. 13–54 A trapezoidal channel with a bottom width of 5 m, free surface width of 10 m, and flow depth of 2.2 m discharges water at a rate of 120 m3/s. If the surfaces of the channel are lined with asphalt (n ! 0.016), determine the elevation drop of the channel per km. Answer: 8.52 m 10 m
2.2 m
5m
FIGURE P13–54 13–55 Reconsider Prob. 13–54. If the maximum flow height the channel can accommodate is 2.4 m, determine the maximum flow rate through the channel. 13–56 Consider water flow through two identical channels with square flow sections of 3 m $ 3 m. Now the two channels are combined, forming a 6-m-wide channel. The flow rate is adjusted so that the flow depth remains constant at 3 m. Determine the percent increase in flow rate as a result of combining the channels.
3m
3m
3m
3m
FIGURE P13–56 13–57 A trapezoidal channel made of unfinished concrete has a bottom slope of 1°, base width of 5 m, and a side surface slope of 1:1, as shown in Fig. P13–57. For a flow rate of 25 m3/s, determine the normal depth h.
y ! 0.45 m u ! 50° b ! 0.6 m
FIGURE P13–51 13–52 Water flows uniformly half-full in a 2-m-diameter circular channel that is laid on a grade of 1.5 m/km. If the channel is made of finished concrete, determine the flow rate of the water.
h 45°
45° 5m
FIGURE P13–57
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13–58 Repeat Prob. 13–57 for a weedy excavated earth channel with n ! 0.030.
For a flow depth of 0.25 m at the center, determine the flow rate of water through the channel.
13–59 A cast iron V-shaped water channel shown in Fig. P13–59 has a bottom slope of 0.5°. For a flow depth of 1 m at the center, determine the discharge rate in uniform flow.
R ! 0.5 m
Answer: 3.59 m3/s
y ! 0.25 m 1m 45°
45°
FIGURE P13–62
FIGURE P13–59
13–63
13–60E Water is to be transported in a cast iron rectangular channel with a bottom width of 6 ft at a rate of 70 ft3/s. The terrain is such that the channel bottom drops 1.5 ft per 1000 ft length. Determine the minimum height of the channel under uniform-flow conditions.
Reconsider Prob. 13–62. By varying the flow depth-to-radius ratio y/R from 0.1 to 1.9 while holding the flow area constant and evaluating the flow rate, show that the best cross section for flow through a circular channel occurs when the channel is half-full. Tabulate and plot your results.
13–64 A clean-earth trapezoidal channel with a bottom width of 1.5 m and a side surface slope of 1:1 is to drain water uniformly at a rate of 8 m3/s to a distance of 1 km. If the flow depth is not to exceed 1 m, determine the required elevation drop.
. V ! 70 ft3/s
13–65 A water draining system with a constant slope of 0.0015 is to be built of three circular channels made of finished concrete. Two of the channels have a diameter of 1.2 m and drain into the third channel. If all channels are to run half-full and the losses at the junction are negligible, determine the diameter of the third channel. Answer: 1.56 m
y
b ! 6 ft
FIGURE P13–60E 13–61 Water flows in a channel whose bottom slope is 0.002 and whose cross section is as shown in Fig. P13–61. The dimensions and the Manning coefficients for the surfaces of different subsections are also given on the figure. Determine the flow rate through the channel and the effective Manning coefficient for the channel. 6m
10 m
1
2
13–66 Water is to be transported in an open channel whose surfaces are asphalt lined at a rate of 4 m3/s in uniform flow. The bottom slope is 0.0015. Determine the dimensions of the best cross section if the shape of the channel is (a) circular of diameter D, (b) rectangular of bottom width b, and (c) trapezoidal of bottom width b. 13–67E A rectangular channel with a bottom slope of 0.0005 is to be built to transport water at a rate of 800 ft3/s. Determine the best dimensions of the channel if it is to be made of (a) unfinished concrete and (b) finished concrete. 13–68
2m
1.5 m
Concrete channel n 1 ! 0.014
Light brush n 2 ! 0.050
2m
Consider uniform flow in an asphalt-lined rectangular channel with a flow area of 2 m2 and a bottom slope of 0.0003. By varying the depth-to-width ratio y/b from 0.1 to 2.0, calculate and plot the flow rate, and confirm that the best flow cross section occurs when the flow depth-to-width ratio is 0.5.
Gradually and Rapidly Varied Flows and Hydraulic Jump
FIGURE P13–61
13–69C How does nonuniform or varied flow differ from uniform flow?
13–62 Consider a 1-m-internal-diameter water channel made of finished concrete (n ! 0.012). The channel slope is 0.002.
13–70C How does gradually varied flow (GVF) differ from rapidly varied flow (RVF)?
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13–71C Someone claims that frictional losses associated with wall shear on surfaces can be neglected in the analysis of rapidly varied flow, but should be considered in the analysis of gradually varied flow. Do you agree with this claim? Justify your answer. 13–72C Consider steady flow of water in a horizontal channel of rectangular cross section. If the flow is subcritical, the flow depth will (a) increase, (b) remain constant, or (c) decrease in the flow direction. 13–73C Consider steady flow of water in a downwardsloped channel of rectangular cross section. If the flow is subcritical and the flow depth is greater than the normal depth (y ' yn), the flow depth will (a) increase, (b) remain constant, or (c) decrease in the flow direction. 13–74C Consider steady flow of water in a horizontal channel of rectangular cross section. If the flow is supercritical, the flow depth will (a) increase, (b) remain constant, or (c) decrease in the flow direction. 13–75C Consider steady flow of water in a downwardsloped channel of rectangular cross section. If the flow is subcritical and the flow depth is less than the normal depth (y ( yn), the flow depth will (a) increase, (b) remain constant, or (c) decrease in the flow direction. 13–76C Consider steady flow of water in an upward-sloped channel of rectangular cross section. If the flow is supercritical, the flow depth will (a) increase, (b) remain constant, or (c) decrease in the flow direction. 13–77C Is it possible for subcritical flow to undergo a hydraulic jump? Explain. 13–78C Why is the hydraulic jump sometimes used to dissipate mechanical energy? How is the energy dissipation ratio for a hydraulic jump defined? 13–79 Water flows uniformly in a rectangular channel with finished-concrete surfaces. The channel width is 3 m, the flow depth is 1.2 m, and the bottom slope is 0.002. Determine if the channel should be classified as mild, critical, or steep for this flow.
13–80 Consider uniform water flow in a wide brick channel of slope 0.4°. Determine the range of flow depth for which the channel is classified as being steep. 13–81E Consider the flow of water through a 12-ft-wide unfinished-concrete rectangular channel with a bottom slope of 0.5°. If the flow rate is 300 ft3/s, determine if the slope of this channel is mild, critical, or steep. Also, for a flow depth of 3 ft, classify the surface profile while the flow develops. 13–82 Water is flowing in a 90° V-shaped cast iron channel with a bottom slope of 0.002 at a rate of 3 m3/s. Determine if the slope of this channel should be classified as mild, critical, or steep for this flow. Answer: mild 13–83
Water discharging into an 8-m-wide rectangular horizontal channel from a sluice gate is observed to have undergone a hydraulic jump. The flow depth and velocity before the jump are 1.2 m and 9 m/s, respectively. Determine (a) the flow depth and the Froude number after the jump, (b) the head loss and the dissipation ratio, and (c) the mechanical energy dissipated by the hydraulic jump.
V1 ! 9 m/s y 1 ! 1.2 m (1)
y2
V2
(2)
FIGURE P13–83 13–84 Water flowing in a wide horizontal channel at a flow depth of 35 cm and an average velocity of 12 m/s undergoes a hydraulic jump. Determine the head loss associated with hydraulic jump. 13–85 During a hydraulic jump in a wide channel, the flow depth increases from 0.6 to 3 m. Determine the velocities and Froude numbers before and after the jump, and the energy dissipation ratio. 13–86 Consider the flow of water in a 10-m-wide channel at a rate of 70 m3/s and a flow depth of 0.50 m. The water now undergoes a hydraulic jump, and the flow depth after the jump is measured to be 4 m. Determine the mechanical power wasted during this jump. Answer: 4.35 MW
y ! 1.2 m
13–87 The flow depth and velocity of water after undergoing a hydraulic jump are measured to be 2 m and 3 m/s, respectively. Determine the flow depth and velocity before the jump, and the fraction of mechanical energy dissipated.
b!3m
13–88E Water flowing in a wide channel at a depth of 2 ft and a velocity of 40 ft/s undergoes a hydraulic jump. Determine the flow depth, velocity, and Froude number after the jump, and the head loss associated with the jump.
FIGURE P13–79
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Flow Control and Measurement in Channels
Depression over the bump
13–89C Draw a flow depth-specific energy diagram for flow through underwater gates, and indicate the flow through the gate for cases of (a) frictionless gate, (b) sluice gate with free outflow, and (c) sluice gate with drowned outflow (including the hydraulic jump back to subcritical flow). 13–90C For sluice gates, how is the discharge coefficient Cd defined? What are typical values of Cd for sluice gates with free outflow? What is the value of Cd for the idealized frictionless flow through the gate? 13–91C What is the basic principle of operation of a broadcrested weir used to measure flow rate through an open channel? 13–92C Consider steady frictionless flow over a bump of height /z in a horizontal channel of constant width b. Will the flow depth y increase, decrease, or remain constant as the fluid flows over the bump? Assume the flow to be subcritical. 13–93C Consider the flow of a liquid over a bump during subcritical flow in an open channel. The specific energy and the flow depth decrease over the bump as the bump height is increased. What will the character of flow be when the specific energy reaches its minimum value? Will the flow become supercritical if the bump height is increased even further? 13–94C What is a sharp-crested weir? On what basis are the sharp-crested weirs classified? 13–95 Water is released from a 14-m-deep reservoir into a 5-m-wide open channel through a sluice gate with a 1-m-high opening at the channel bottom. If the flow depth downstream from the gate is measured to be 3 m, determine the rate of discharge through the gate.
y1 ! 1.2 m
V1 ! 2.5 m/s
y2
/zb ! 0.22 m
Bump
FIGURE P13–96 13–97 Consider the uniform flow of water in a wide channel with a velocity of 8 m/s and flow depth of 0.8 m. Now water flows over a 30-cm-high bump. Determine the change (increase or decrease) in the water surface level over the bump. Also determine if the flow over the bump is sub- or supercritical. 13–98 The flow rate of water in a 4-m-wide horizontal channel is being measured using a 0.75-m-high sharp-crested rectangular weir that spans across the channel. If the water depth upstream is 2.2 m, determine the flow rate of water. Answer: 15.9 m3/s
y1 ! 2.2 m V1 Pw ! 0.75 m
Sharp-crested rectangular weir
FIGURE P13–98
Sluice gate
13–99 Repeat Prob. 13–98 for the case of a weir height of 1 m.
y 1 ! 14 m y2 ! 3 m a!1m
FIGURE P13–95 13–96 Water flowing in a wide channel encounters a 22cm-high bump at the bottom of the channel. If the flow depth is 1.2 m and the velocity is 2.5 m/s before the bump, determine if the flow is chocked over the bump, and discuss.
13–100 Water flows over a 2-m-high sharp-crested rectangular weir. The flow depth upstream of the weir is 3 m, and water is discharged from the weir into an unfinished-concrete channel of equal width where uniform-flow conditions are established. If no hydraulic jump is to occur in the downstream flow, determine the maximum slope of the downstream channel. 13–101E A full-width sharp-crested weir is to be used to measure the flow rate of water in a 10-ft-wide rectangular channel. The maximum flow rate through the channel is 150 ft3/s, and the flow depth upstream from the weir is not to exceed 5 ft. Determine the appropriate height of the weir. 13–102 Consider uniform water flow in a wide rectangular channel with a depth of 2 m made of unfinished concrete laid on a slope of 0.0022. Determine the flow rate of water per m
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731 CHAPTER 13
width of channel. Now water flows over a 15-cm-high bump. If the water surface over the bump remains flat (no rise or drop), determine the change in discharge rate of water per meter width of the channel. Hint: Investigate if a flat surface over the bump is physically possible. 13–103 Consider uniform water flow in a wide channel made of unfinished concrete laid on a slope of 0.0022. Now water flows over a 15-cm-high bump. If the flow over the bump is exactly critical (Fr ! 1), determine the flow rate and the flow depth over the bump per m width. Answers: 20.3 m3/s, 3.48 m
y1
13–108E Consider water flow through a wide channel at a flow depth of 8 ft. Now water flows through a sluice gate with a 1-ft-high opening, and the freely discharged outflow subsequently undergoes a hydraulic jump. Disregarding any losses associated with the sluice gate itself, determine the flow depth and velocities before and after the jump, and the fraction of mechanical energy dissipated during the jump. 13–109 The flow rate of water flowing in a 3-m-wide channel is to be measured with a sharp-crested triangular weir 0.5 m above the channel bottom with a notch angle of 60°. If the flow depth upstream from the weir is 1.5 m, determine the flow rate of water through the channel. Take the weir discharge coefficient to be 0.60. Answer: 0.818 m3/s
y2
Upstream free surface
/zb ! 15 cm
Bump
1m
Slope ! 0.0022
FIGURE P13–103
60° Weir plate
0.5 m
13–104 The flow rate of water through a 5-m-wide (into the paper) channel is controlled by a sluice gate. If the flow depths are measured to be 1.1 and 0.45 m upstream and downstream from the gates, respectively, determine the flow rate and the Froude number downstream from the gate. Sluice gate
3m
FIGURE P13–109 13–110 1.2 m.
Repeat Prob. 13–109 for an upstream flow depth of
13–111 A sharp-crested triangular weir with a notch angle of 100° is used to measure the discharge rate of water from a large lake into a spillway. If a weir with half the notch angle (u ! 50°) is used instead, determine the percent reduction in the flow rate. Assume the water depth in the lake and the weir discharge coefficient remain unchanged.
y 1 ! 1.1 m
y 2 ! 0.45 m
FIGURE P13–104
13–112 A 1-m-high broad-crested weir is used to measure the flow rate of water in a 5-m-wide rectangular channel. The flow depth well upstream from the weir is 1.6 m. Determine the flow rate through the channel and the minimum flow depth above the weir.
13–105E Water flows through a sluice gate with a 1.1-fthigh opening and is discharged with free outflow. If the upstream flow depth is 5 ft, determine the flow rate per unit width and the Froude number downstream the gate.
Discharge 1.6 m 1m
13–106E Repeat Prob. 13–105E for the case of a drowned gate with a downstream flow depth of 3.3 ft. 13–107 Water is to be discharged from a 6-m-deep lake into a channel through a sluice gate with a 5-m wide and 0.6m-high opening at the bottom. If the flow depth downstream from the gate is measured to be 3 m, determine the rate of discharge through the gate.
Broad-crested weir
FIGURE P13–112 13–113 2.2 m.
Repeat Prob. 13–112 for an upstream flow depth of
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732 FLUID MECHANICS
13–114 Consider water flow over a 0.80-m-high sufficiently long broad-crested weir. If the minimum flow depth above the weir is measured to be 0.50 m, determine the flow rate per meter width of channel and the flow depth upstream of the weir.
R!1m 1.5 m
Review Problems 13–115 A trapezoidal channel with a bottom width of 4 m and a side slope of 45° discharges water at a rate of 18 m3/s. If the flow depth is 0.6 m, determine if the flow is subcritical or supercritical. 13–116 A rectangular channel with a bottom width of 2 m discharges water at a rate of 8 m3/s. Determine the flow depth below which the flow is supercritical.
FIGURE P13–122 13–123 Consider water flow through a V-shaped channel. Determine the angle u the channel makes from the horizontal for which the flow is most efficient.
13–117 Water flows in a canal at an average velocity of 4 m/s. Determine if the flow is subcritical or supercritical for flow depths of (a) 0.2 m, (b) 2 m, and (c) 1.63 m. 13–118 Water flows through a 1.5-m-wide rectangular channel with a Manning coefficient of n ! 0.012. If the water is 0.9 m deep and the bottom slope of the channel is 0.6°, determine the rate of discharge of the channel in uniform flow. 13–119
A 5-m-wide rectangular channel lined with finished concrete is to be designed to transport water to a distance of 1 km at a rate of 12 m3/s. Using EES (or other) software, investigate the effect of bottom slope on flow depth (and thus on the required channel height). Let the bottom angle vary from 0.5 to 10° in increments of 0.5°. Tabulate and plot the flow depth against the bottom angle, and discuss the results. 13–120
Repeat Prob. 13–119 for a trapezoidal channel that has a base width of 5 m and a side surface angle of 45°. 13–121 A trapezoidal channel with brick lining has a bottom slope of 0.001 and a base width of 4 m, and the side surfaces are angled 30° from the horizontal, as shown in Fig. P13–121. If the normal depth is measured to be 2 m, estimate the flow rate of water through the channel. Answer: 36.4 m3/s
y u
u
FIGURE P13–123 13–124E A rectangular channel with unfinished concrete surfaces is to be built to discharge water uniformly at a rate of 200 ft3/s. For the case of best cross section, determine the bottom width of the channel if the available vertical drop is (a) 8 and (b) 10 ft per mile. Answers: (a) 7.86 and (b) 7.54 ft per mile
13–125E Repeat Prob. 13–124E for the case of a trapezoidal channel of best cross section. 13–126 Water flows in a channel whose bottom slope is 0.5° and whose cross section is as shown in Fig. P13–126. The dimensions and the Manning coefficients for the surfaces of different subsections are also given on the figure. Determine the flow rate through the channel and the effective Manning coefficient for the channel. 6m 1m
1m
Clean earth channel n 1 ! 0.022
10 m
Heavy brush n 2 ! 0.075
2m 30°
30° 4m
FIGURE P13–121 13–122 A 2-m-internal-diameter circular steel storm drain (n ! 0.012) is to discharge water uniformly at a rate of 12 m3/s to a distance of 1 km. If the maximum depth is to be 1.5 m, determine the required elevation drop.
3m
FIGURE P13–126 13–127 Consider two identical channels, one rectangular of bottom width b and one circular of diameter D, with identical flow rates, bottom slopes, and surface linings. If the flow height in the rectangular channel is also b and the circular channel is flowing half-full, determine the relation between b and D.
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733 CHAPTER 13
13–128 Consider the flow of water through a parabolic notch shown in Fig. P13–128. Develop a relation for the flow rate, and calculate its numerical value for the ideal case in which the flow velocity is given by Toricelli’s equation V ! 12g(H % y). Answer: 0.246 m3/s b ! 0.4 m y ! cx2
H ! 0.5 m
y x
13–133 Consider water flow through a wide rectangular channel undergoing a hydraulic jump. Show that the ratio of the Froude numbers before and after the jump can be expressed in terms of flow depths y1 and y2 before and after the jump, respectively, as Fr1/Fr2 ! 1(y 2,y 1)3. 13–134 A sluice gate with free outflow is used to control the discharge rate of water through a channel. Determine the flow rate per unit width when the gate is raised to yield a gap of 30 cm and the upstream flow depth is measured to be 1.8 m. Also determine the flow depth and the velocity downstream. 13–135 Water flowing in a wide channel at a flow depth of 45 cm and an average velocity of 8 m/s undergoes a hydraulic jump. Determine the fraction of the mechanical energy of the fluid dissipated during this jump. Answer: 36.8 percent
FIGURE P13–128 13–129
In practice, the V-notch is commonly used to measure flow rate in open channels. Using the idealized Toricelli’s equation V ! 12g(H % y) for velocity, develop a relation for the flow rate through the V-notch in terms of the angle u. Also, show the variation of the flow rate with u by evaluating the flow rate for u ! 25, 40, 60, and 75°, and plotting the results.
13–136 Water flowing through a sluice gate undergoes a hydraulic jump, as shown in Fig. P13–136. The velocity of the water is 1.25 m/s before reaching the gate and 4 m/s after the jump. Determine the flow rate of water through the gate per meter of width, the flow depths y1 and y2, and the energy dissipation ratio of the jump. Sluice gate
H ! 25 cm u
y1
y
FIGURE P13–129
y2
13–130 Water flows uniformly half-full in a 1.2-m-diameter circular channel laid with a slope of 0.004. If the flow rate of water is measured to be 1.25 m3/s, determine the Manning coefficient of the channel and the Froude number. 13–131 Water flowing in a wide horizontal channel approaches a 20-cm-high bump with a velocity of 1.25 m/s and a flow depth of 1.8 m. Determine the velocity, flow depth, and Froude number over the bump.
y1 ! 1.8 m V1 ! 1.25 m/s
y2
V2 20 cm
FIGURE P13–131 13–132 Reconsider Prob. 13–131. Determine the bump height for which the flow over the bump is critical (Fr ! 1).
y3 ! 3 m V3 ! 4 m/s
V1 ! 1.25 m/s
FIGURE P13–136 13–137 Repeat Prob. 13–136 for a velocity of 2 m/s after the hydraulic jump. 13–138 Water is discharged from a 5-m-deep lake into a finished concrete channel with a bottom slope of 0.004 through a sluice gate with a 0.5-m-high opening at the bottom. Shortly after supercritical uniform-flow conditions are established, the water undergoes a hydraulic jump. Determine the flow depth, velocity, and Froude number after the jump. Disregard the bottom slope when analyzing the hydraulic jump. 13–139 Water is discharged from a dam into a wide spillway to avoid overflow and to reduce the risk of flooding. A large fraction of the destructive power of water is dissipated by a hydraulic jump during which the water depth rises from
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734 FLUID MECHANICS
0.50 to 4 m. Determine the velocities of water before and after the jump, and the mechanical power dissipated per meter width of the spillway. 13–140 The flow rate of water in a 6-m-wide rectangular channel is to be measured using a 1.1-m-high sharp-crested rectangular weir that spans across the channel. If the head above the weir crest is 0.60 m upstream from the weir, determine the flow rate of water. 13–141E Consider two identical 12-ft-wide rectangular channels each equipped with a 2-ft-high full-width weir, except that the weir is sharp-crested in one channel and broad-crested in the other. For a flow depth of 5 ft in both channels, determine the flow rate through each channel. Answers: 244 ft3/s, 79.2 ft3/s
Design and Essay Problems 13–142 Using catalogs or websites, obtain information from three different weir manufacturers. Compare the different weir designs, and discuss the advantages and disadvantages of each design. Indicate the applications for which each design is best suited. 13–143 Consider water flow in the range of 10 to 15 m3/s through a horizontal section of a 5-m-wide rectangular channel. A rectangular or triangular thin-plate weir is to be installed to measure the flow rate. If the water depth is to remain under 2 m at all times, specify the type and dimensions of an appropriate weir. What would your response be if the flow range were 0 to 15 m3/s?
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CHAPTER
14
TURBOMACHINERY
I
n this chapter we discuss the basic principles of a common and important application of fluid mechanics, turbomachinery. First we classify turbomachines into two broad categories, pumps and turbines. Then we discuss both of these turbomachines in more detail, mostly qualitatively, explaining the basic principles of their operation. We emphasize preliminary design and overall performance of turbomachines rather than detailed design. In addition, we discuss how to properly match the requirements of a fluid flow system to the performance characteristics of a turbomachine. A significant portion of this chapter is devoted to turbomachinery scaling laws—a practical application of dimensional analysis. We show how the scaling laws are used in the design of new turbomachines that are geometrically similar to existing ones.
OBJECTIVES When you finish reading this chapter, you should be able to ■
■
■
■
Identify various types of pumps and turbines, and understand how they work Apply dimensional analysis to design new pumps or turbines that are geometrically similar to existing pumps or turbines Perform basic vector analysis of the flow into and out of pumps and turbines Use specific speed for preliminary design and selection of pumps and turbines
735
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736 FLUID MECHANICS
14–1 Flow in Ein
Pump
Flow out Eout
v Energy supplied, Eout > Ein (a)
Flow in Ein
Turbine
Flow out Eout
v Energy extracted, Eout < Ein (b)
FIGURE 14–1 (a) A pump supplies energy to a fluid, while (b) a turbine extracts energy from a fluid.
Control volume
Din Vin
Pump
Pin
Dout Vout Pout
v
FIGURE 14–2 For the case of steady flow, conservation of mass requires that the mass flow rate out of a pump must equal the mass flow rate into the pump; for incompressible flow with equal inlet and outlet cross-sectional areas (Dout ! Din), we conclude that Vout ! Vin, but Pout " Pin.
∆P •
V
Fan Low
Blower Medium
Compressor High
High
Medium
Low
FIGURE 14–3 When used with gases, pumps are called fans, blowers, or compressors, depending on the relative values of pressure rise and volume flow rate.
■
CLASSIFICATIONS AND TERMINOLOGY
There are two broad categories of turbomachinery, pumps and turbines. The word pump is a general term for any fluid machine that adds energy to a fluid. Some authors call pumps energy absorbing devices since energy is supplied to them, and they transfer most of that energy to the fluid, usually via a rotating shaft (Fig. 14–1a). The increase in fluid energy is usually felt as an increase in the pressure of the fluid. Turbines, on the other hand, are energy producing devices—they extract energy from the fluid and transfer most of that energy to some form of mechanical energy output, typically in the form of a rotating shaft (Fig. 14–1b). The fluid at the outlet of a turbine suffers an energy loss, typically in the form of a loss of pressure. An ordinary person may think that the energy supplied to a pump increases the speed of fluid passing through the pump and that a turbine extracts energy from the fluid by slowing it down. This is not necessarily the case. Consider a control volume surrounding a pump (Fig. 14–2). We assume steady conditions. By this we mean that neither the mass flow rate nor the rotational speed of the rotating blades changes with time. (The detailed flow field near the rotating blades inside the pump is not steady of course, but control volume analysis is not concerned with details inside the control volume.) By conservation of mass, we know that the mass flow rate into the pump must equal the mass flow rate out of the pump. If the flow is incompressible, the volume flow rates at the inlet and outlet must be equal as well. Furthermore, if the diameter of the outlet is the same as that of the inlet, conservation of mass requires that the average speed across the outlet must be identical to the average speed across the inlet. In other words, the pump does not necessarily increase the speed of the fluid passing through it; rather, it increases the pressure of the fluid. Of course, if the pump were turned off, there might be no flow at all. So, the pump does increase fluid speed compared to the case of no pump in the system. However, in terms of changes from the inlet to the outlet across the pump, fluid speed is not necessarily increased. (The output speed may even be lower than the input speed if the outlet diameter is larger than that of the inlet.) The purpose of a pump is to add energy to a fluid, resulting in an increase in fluid pressure, not necessarily an increase of fluid speed across the pump.
An analogous statement is made about the purpose of a turbine: The purpose of a turbine is to extract energy from a fluid, resulting in a decrease of fluid pressure, not necessarily a decrease of fluid speed across the turbine.
Fluid machines that move liquids are called pumps, but there are several other names for machines that move gases (Fig. 14–3). A fan is a gas pump with relatively low pressure rise and high flow rate. Examples include ceiling fans, house fans, and propellers. A blower is a gas pump with relatively moderate to high pressure rise and moderate to high flow rate. Examples include centrifugal blowers and squirrel cage blowers in automobile ventilation systems, furnaces, and leaf blowers. A compressor is a gas pump designed to deliver a very high pressure rise, typically at low to moderate flow rates. Examples include air compressors that run pneumatic tools and inflate tires at automobile service stations, and refrigerant compressors used in heat pumps, refrigerators, and air conditioners.
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737 CHAPTER 14
Pumps and turbines in which energy is supplied or extracted by a rotating shaft are properly called turbomachines, since the Latin prefix turbo means “spin.” Not all pumps or turbines utilize a rotating shaft, however. The hand-operated air pump you use to inflate the tires of your bicycle is a prime example (Fig. 14–4a). The up and down reciprocating motion of a plunger or piston replaces the rotating shaft in this type of pump, and it is more proper to call it simply a fluid machine instead of a turbomachine. An old-fashioned well pump operates in a similar manner to pump water instead of air (Fig. 14–4b). Nevertheless, the words turbomachine and turbomachinery are often used in the literature to refer to all types of pumps and turbines regardless of whether they utilize a rotating shaft or not. Fluid machines may also be broadly classified as either positive-displacement machines or dynamic machines, based on the manner in which energy transfer occurs. In positive-displacement machines, fluid is directed into a closed volume. Energy transfer to the fluid is accomplished by movement of the boundary of the closed volume, causing the volume to expand or contract, thereby sucking fluid in or squeezing fluid out, respectively. Your heart is a good example of a positive-displacement pump (Fig. 14–5a). It is designed with one-way valves that open to let blood in as heart chambers expand, and other one-way valves that open as blood is pushed out of those chambers when they contract. An example of a positive-displacement turbine is the common water meter in your house (Fig. 14–5b), in which water forces itself into a closed chamber of expanding volume connected to an
Superior vena cava
Pulmonary artery
Aorta
Left atrium Right atrium Left ventricle
Pulmonary valve
Pulmonary vein Mitral valve Aortic valve
Right ventricle Inferior vena cava
Tricuspid valve
(a)
(b)
FIGURE 14–5 (a) The human heart is an example of a positive-displacement pump; blood is pumped by expansion and contraction of heart chambers called ventricles. (b) The common water meter in your house is an example of a positive-displacement turbine; water fills and exits a chamber of known volume for each revolution of the output shaft. Photo courtesy of Niagara Meters, Spartanburg, SC. Used by permission.
(a)
(b)
FIGURE 14–4 Not all pumps have a rotating shaft; (a) energy is supplied to this manual tire pump by the up and down motion of a person’s arm to pump air; (b) a similar mechanism is used to pump water with an old-fashioned well pump. (a) Photo by Andrew Cimbala, with permission. (b) © The McGraw-Hill Companies, Inc./Ellen Behrman, photographer.
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738 FLUID MECHANICS
output shaft that turns as water enters the chamber. The boundary of the volume then collapses, turning the output shaft some more, and letting the water continue on its way to your sink, shower, etc. The water meter records each 360° rotation of the output shaft, and the meter is precisely calibrated to the known volume of fluid in the chamber. In dynamic machines, there is no closed volume; instead, rotating blades supply or extract energy to or from the fluid. For pumps, these rotating blades are called impeller blades, while for turbines, the rotating blades are called runner blades or buckets. Examples of dynamic pumps include enclosed pumps and ducted pumps (those with casings around the blades such as the water pump in your car’s engine), and open pumps (those without casings such as the ceiling fan in your house, the propeller on an airplane, or the rotor on a helicopter). Examples of dynamic turbines include enclosed turbines, such as the hydroturbine that extracts energy from water in a hydroelectric dam, and open turbines such as the wind turbine that extracts energy from the wind (Fig. 14–6).
FIGURE 14–6 A wind turbine is a good example of a dynamic machine of the open type; air turns the blades, and the output shaft drives an electric generator. The Wind Turbine Company. Used by permission.
14–2
■
PUMPS
Some fundamental parameters are used to analyze the performance of a . pump. The mass flow rate of fluid through the pump, m, is an obvious primary pump performance parameter. For incompressible flow, it is more common to use volume flow rate rather than mass flow rate. In the turbomachinery industry, volume flow rate is called capacity and is simply mass flow rate divided by fluid density, Volume flow rate (capacity):
# # m V! r
(14–1)
The performance of a pump is characterized additionally by its net head H, defined as the change in Bernoulli head between the inlet and outlet of the pump, Net head:
H! a
V2 V2 P P $ $ $ zb # a $ zb rg 2g rg 2g out in
(14–2)
The dimension of net head is length, and it is often listed as an equivalent column height of water, even for a pump that is not pumping water. For the case in which a liquid is being pumped, the Bernoulli head at the inlet is equivalent to the energy grade line at the inlet, EGLin, obtained by aligning a Pitot probe in the center of the flow as illustrated in Fig. 14–7. The energy grade line at the outlet EGLout is obtained in the same manner, as also illustrated in the figure. In the general case, the outlet of the pump may be at a different elevation than the inlet, and its diameter and average speed may not be the same as those at the inlet. Regardless of these differences, net head H is equal to the difference between EGLout and EGLin, Net head for a liquid pump:
H ! EGLout # EGLin
Consider the special case of incompressible flow through a pump in which the inlet and outlet diameters are identical, and there is no change in elevation. Equation 14–2 reduces to Special case with Dout ! Din and zout ! zin:
H!
Pout # Pin rg
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739 CHAPTER 14
For this simplified case, net head is simply the pressure rise across the pump expressed as a head (column height of the fluid). Net head is proportional to the useful power actually delivered to the fluid. It is traditional to call this power the water horsepower, even if the fluid being pumped is not water, and even if the power is not measured in units of horsepower. By dimensional reasoning, we must multiply the net head of Eq. 14–2 by mass flow rate and gravitational acceleration to obtain dimensions of power. Thus, Water horsepower:
# # # Wwater horsepower ! mgH ! rgV H
Brake horsepower:
(14–4)
where v is the rotational speed of the shaft (rad/s) and Tshaft is the torque supplied to the shaft. We define pump efficiency hpump as the ratio of useful power to supplied power, Pump efficiency:
# # # Wwater horsepower Wwater horsepower rgV H h pump ! ! ! # bhp vTshaft Wshaft
Pump Performance Curves and Matching a Pump to a Piping System
zout
(14–3)
All pumps suffer from irreversible losses due to friction, internal leakage, flow separation on blade surfaces, turbulent dissipation, etc. Therefore, the . mechanical energy supplied to the pump must be larger than Wwater horsepower. In pump terminology, the external power supplied to the pump is called the brake horsepower, which we abbreviate as bhp. For the typical case of a rotating shaft supplying the brake horsepower, # bhp ! Wshaft ! vTshaft
H
(14–5)
The maximum volume flow rate through a pump occurs when its net head is zero, H ! 0; this flow rate is called the pump’s free delivery. The free delivery condition is achieved when there is no flow restriction at the pump inlet or outlet—in. other words when there is no load on the pump. At this operating point, V is large, but H is zero; the pump’s efficiency is zero because the pump is doing no useful work, as is clear from Eq. 14–5. At the other extreme, the .shutoff head is the net head that occurs when the volume flow rate is zero, V ! 0, and is achieved when the outlet . port of the pump is blocked off. Under these conditions, H is large but V is zero; the pump’s efficiency (Eq. 14–5) is again zero, because the pump is doing no useful work. Between these two extremes, from shutoff to free delivery, the pump’s net head may increase from its shutoff value somewhat as the flow rate increases, but H must eventually decrease to zero as the volume flow rate increases to its free delivery value. The pump’s efficiency reaches its maximum value somewhere between the shutoff condition and the free delivery condition; this operating point of maximum efficiency is appropriately called the best efficiency point (BEP), and is notated by .an asterisk . (H*, V *, bhp*). Curves of H, hpump, and bhp as functions of V are called pump performance curves (or characteristic curves, Chap. 8); typical curves at one rotational speed are plotted in Fig. 14–8. The pump performance curves change with rotational speed. It is important to realize that for steady conditions, a pump can operate only along its performance curve. Thus, the operating point of a piping system is
Din Vin
Pump
Vout Pout Dout
zin
Pin
EGLin
v bhp
EGLout
Datum plane (z = 0)
FIGURE 14–7 The net head of a pump, H, is defined as the change in Bernoulli head from inlet to outlet; for a liquid, this is equivalent to the change in the energy grade line, H ! EGLout # EGLin, relative to some arbitrary datum plane; bhp is the brake horsepower, the external power supplied to the pump.
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740 FLUID MECHANICS Shutoff head
determined by matching system requirements (required net head) to pump performance (available net head). In a typical application, Hrequired and Havailable match at one unique value of flow rate—this is the operating point or duty point of the system.
BEP
H, hpump, or bhp
H H*
The steady operating point of a piping system is established at the volume flow rate where Hrequired ! Havailable.
hpump bhp* bhp •
V* 0
•
0
Free delivery
V
FIGURE 14–8 Typical pump performance curves for a centrifugal pump with backwardinclined blades; the curve shapes for other types of pumps may differ, and the curves change as shaft rotation speed is changed. Pump performance curve H Operating point
Havailable
System curve Hrequired 0 0
BEP
•
V
FIGURE 14–9 The operating point of a piping system is established as the volume flow rate where the system curve and the pump performance curve intersect.
For a given piping system with its major and minor losses, elevation changes, etc., the required net head increases with volume flow rate. On the other hand, the available net head of most pumps decreases with flow rate, as in Fig. 14–8, at least over the majority of its recommended operating range. Hence, the system curve and the pump performance curve intersect as sketched in Fig. 14–9, and this establishes the operating point. If we are lucky, the operating point is at or near the best efficiency point of the pump. In most cases, however, as illustrated in Fig. 14–9, the pump does not run at its optimum efficiency. If efficiency is of major concern, the pump should be carefully selected (or a new pump should be designed) such that the operating point is as close to the best efficiency point as possible. In some cases it may be possible to change the shaft rotation speed so that an existing pump can operate much closer to its design point (best efficiency point). There are unfortunate situations where the system curve and the pump performance curve intersect at more than one operating point. This can occur when a pump that has a dip in its net head performance curve is mated to a system that has a fairly flat system curve, as illustrated in Fig. 14-10. Although rare, such situations are possible and should be avoided, because the system may “hunt” for an operating point, leading to an unsteadyflow situation. It is fairly straightforward to match a piping system to a pump, once we realize that the term for useful pump head (hpump, u) that we used in the head form of the energy equation (Chap. 5) is the same as the net head (H) used in the present chapter. Consider, for example, a general piping system with elevation change, major and minor losses, and fluid acceleration (Fig. 14–11). We begin by solving the energy equation for the required net head Hrequired, Hrequired ! hpump, u !
P2 # P1 a 2V 22 # a 1V 21 $ $ (z 2 # z 1) $ hturbine $ hL, total (14–6) rg 2g
where we assume that there is no turbine in the system, although that term can be added back in, if necessary. We have also included the kinetic energy correction factors in Eq. 14–6 for greater accuracy, even though it is common practice in the turbomachinery industry to ignore them (a1 and a2 are often assumed to be unity since the flow is turbulent). Equation 14–6 is evaluated from the inlet of the piping system (point 1, upstream of the pump) to the outlet of the piping system (point 2, downstream of the pump). Equation 14–6 agrees with our intuition, because it tells us that the useful pump head delivered to the fluid does four things: • It increases the static pressure of the fluid from point 1 to point 2 (first term on the right). • It increases the dynamic pressure (kinetic energy) of the fluid from point 1 to point 2 (second term on the right).
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741 CHAPTER 14
• It raises the elevation (potential energy) of the fluid from point 1 to point 2 (third term on the right). • It overcomes irreversible head losses in the piping system (last term on the right). In a general system, the change in static pressure, dynamic pressure, and elevation may be either positive or negative, while irreversible head losses are always positive. In many mechanical and civil engineering problems in which the fluid is a liquid, the elevation term is important, but when the fluid is a gas, such as in ventilation and air pollution control problems, the elevation term is almost always negligible. To match a pump to a system, and to determine the operating point, we equate Hrequired of Eq. 14–6 to Havailable, which is the (typically known) net head of the pump as a function of volume flow rate. Operating point:
Hrequired ! Havailable
H H available
Hrequired Possible operating points
0 0
•
V
FIGURE 14–10 Situations in which there can be more than one unique operating point should be avoided. In such cases a different pump should be used.
(14–7) z2
The most common situation is that an engineer selects a pump that is somewhat heftier than actually required. The volume flow rate through the piping system is then a bit larger than needed, and a valve or damper is installed in the line so that the flow rate can be decreased as necessary.
V2 2 z2 – z1 1
EXAMPLE 14–1
Operating Point of a Fan in a Ventilation System
A local ventilation system (hood and exhaust duct) is used to remove air and contaminants produced by a dry-cleaning operation (Fig. 14–12). The duct is round and is constructed of galvanized steel with longitudinal seams and with joints every 30 in (0.76 m). The inner diameter (ID) of the duct is D ! 9.06 in (0.230 m), and its total length is L ! 44.0 ft (13.4 m). There are five CD3-9 elbows along the duct. The equivalent roughness height of this duct is 0.15 mm, and each elbow has a minor (local) loss coefficient of KL ! C0 ! 0.21. Note the notation C0 for minor loss coefficient, commonly used in the ventilation industry (ASHRAE, 2001). To ensure adequate venti-. lation, the minimum required volume flow rate through the duct is V ! 600 cfm (cubic feet per minute), or 0.283 m3/s at 25°C. Literature from the hood manufacturer lists the hood entry loss coefficient as 1.3 based on duct velocity. When the damper is fully open, its loss coefficient is 1.8. A centrifugal fan with 9.0-in inlet and outlet diameters is available. Its performance data are shown in Table 14–1, as listed by the manufacturer. Predict the operating point of this local ventilation system, and draw a plot of required and available fan pressure rise as functions of volume flow rate. Is the chosen fan adequate?
SOLUTION We are to estimate the operating point for a given fan and duct system and to plot required and available fan pressure rise as functions of volume flow rate. We are then to determine if the selected fan is adequate. Assumptions 1 The flow is steady. 2 The concentration of contaminants in the air is low; the fluid properties are those of air alone. 3 The flow at the outlet is fully developed turbulent pipe flow with a ! 1.05. Properties For air at 25°C, n ! 1.562 % 10#5 m2/s and r ! 1.184 kg/m3. Standard atmospheric pressure is Patm ! 101.3 kPa.
V1 ! 0
z1
Reservoir
Valve
Pump Valve
FIGURE 14–11 Equation 14–6 emphasizes the role of a pump in a piping system; namely, it increases (or decreases) the static pressure, dynamic pressure, and elevation of the fluid, and it overcomes irreversible losses.
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742 FLUID MECHANICS z2
•
V 2
Analysis We apply the steady energy equation in head form (Eq. 14–6) from point 1 in the stagnant air region in the room to point 2 at the duct outlet,
Hrequired !
(1)
P2 # P1 a 2V 22 # a 1V 21 $ $ (z 2 # z 1) $ hL, total rg 2g negligible for gases
In Eq. 1 we may ignore the air speed at point 1 since it was chosen (wisely) far enough away from the hood inlet so that the air is nearly stagnant. At point 1, P1 is equal to Patm, and at point 2, P2 is also equal to Patm since the jet discharges into the outside air on the roof of the building. Thus the pressure terms cancel out and Eq. 1 reduces to
Fan
Hrequired !
Required net head:
Damper
a 2V 22 $ hL, total 2g
(2)
The total head loss in Eq. 2 is a combination of major and minor losses and depends on volume flow rate. Since the duct diameter is constant, Hood
h L, total ! af
Total irreversible head loss: 1
(3)
The dimensionless roughness factor is e/D ! (0.15 mm)/(230 mm) ! 6.52 % 10#4. The Reynolds number of the air flowing through the duct is
z1
FIGURE 14–12 The local ventilation system for Example 14–1, showing the fan and all minor losses.
# # DV D 4V 4V ! Re ! ! n n pD2 npD
Reynolds number:
Manufacturer’s performance data for the fan of Example 14–1* . V, cfm (dP)fan, inches H2O 0.90 0.95 0.90 0.75 0.40 0.0
* Note that the pressure rise data are listed as inches of water, even though air is the fluid. This is common practice in the ventilation industry.
(4)
The Reynolds number varies with volume flow rate. At the minimum required flow rate, the air speed through the duct is V ! V2 ! 6.81 m/s, and the Reynolds number is
TA B L E 1 4 – 1
0 250 500 750 1000 1200
L V2 $ a K Lb D 2g
Re !
4(0.283 m3/s) ! 1.00 % 10 5 (1.562 % 10 #5 m2/s)p(0.230 m)
From the Moody chart (or the Colebrook equation) at this Reynolds number and roughness factor, the friction factor is f ! 0.0209. The sum of all the minor loss coefficients is
a K L ! 1.3 $ 5(0.21) $ 1.8 ! 4.15
Minor losses:
(5)
Substituting these values at the minimum required flow rate into Eq. 2, the required net head of the fan at the minimum flow rate is
Hrequired ! aa 2 $ f
L V2 $ a K Lb D 2g
! a1.05 $ 0.0209
(6.81 m/s)2 13.4 m $ 4.15b ! 15.2 m of air 0.230 m 2(9.81 m/s2)
(6)
Note that the head is expressed naturally in units of equivalent column height of the pumped fluid, which is air in this case. We convert to an equivalent column height of water by multiplying by the ratio of air density to water density,
cen72367_ch14.qxd 11/12/04 2:11 PM Page 743
743 CHAPTER 14
Hrequired, inches of water ! Hrequired, air
1
r air r water
0.9 3
1.184 kg/m
1 in a b 998.0 kg/m3 0.0254 m
! 0.709 inches of water
0.8
Operating point
0.7 (7)
We repeat the calculations at several values of volume flow rate, and compare to the available net head of the fan in Fig. 14–13. The operating point is at a volume flow rate of about 650 cfm, at which both the required and available net head equal about 0.83 inches of water. We conclude that the chosen fan is more than adequate for the job. Discussion The purchased fan is somewhat more powerful than required, yielding a higher flow rate than necessary. The difference is small and is acceptable; the butterfly damper valve could be partially closed to cut back the flow rate to 600 cfm if necessary. For safety reasons, it is clearly better to oversize than undersize a fan when used with an air pollution control system.
It is common practice in the pump industry to offer several choices of impeller diameter for a single pump casing. There are several reasons for this: (1) to save manufacturing costs, (2) to enable capacity increase by simple impeller replacement, (3) to standardize installation mountings, and (4) to enable reuse of equipment for a different application. When plotting the performance of such a “family” of pumps, pump manufacturers do not plot separate curves of H, hpump, and bhp for each impeller diameter in the form sketched in Fig. 14–8. Instead, they prefer to combine the performance curves of an entire family of pumps of different impeller diameters onto a single plot (Fig. 14–14). Specifically, they plot a curve of H as a function of . V for each impeller diameter in the same way as in Fig. 14–8, but create contour lines of constant efficiency, by drawing smooth curves through points that have the same value of hpump for the various choices of impeller diameter. Contour lines of constant bhp are often drawn on the same plot in similar fashion. An example is provided in Fig. 14–15 for a family of centrifugal pumps manufactured by Taco, Inc. In this case, five impeller diameters are available, but the identical pump casing is used for all five options. As seen in Fig. 14–15, pump manufacturers do not always plot their pumps’ performance curves all the way to free delivery. This is because the pumps are usually not operated there due to the low values of net head and efficiency. If higher values of flow rate and net head are required, the customer should step up to the next larger casing size, or consider using additional pumps in series or parallel. It is clear from the performance plot of Fig. 14–15 that for a given pump casing, the larger the impeller, the higher the maximum achievable efficiency. Why then would anyone buy the smaller impeller pump? To answer this question, we must recognize that the customer’s application requires a certain combination of flow rate and net head. If the requirements match a particular impeller diameter, it may be more cost effective to sacrifice pump efficiency in order to satisfy these requirements.
H, inches H2O
! (15.2 m)
Havailable
0.6 0.5 0.4 0.3
Hrequired
0.2 0.1 0 0
200
400• 600 V, cfm
800
1000
FIGURE 14–13 Net head as a function of volume flow rate for the ventilation system of Example 14–1. The point where the available and required values of H intersect is the operating point.
H
D4 D3 D2 D1
70%
BEP, hpump = 85%
80% 80%
75% 70% 60% 50%
0
•
0
V
FIGURE 14–14 Typical pump performance curves for a family of centrifugal pumps of the same casing diameter but different impeller diameters.
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744 FLUID MECHANICS
5
10
15
20
25
30
35
40
45
50
55
NPSH 60
NPSH REQUIRED
100
kPa
L/s
120
Curve no. 2313 Min. Imp. Bio. 6.75" Size 5 % 4 % 12.75
1160 RPM
Fl & Cl Series
Feet
Model 4013
10 8 6 4 2 0
30 25 20 15 10 5 0
25
250
20
200
15
150
40
8.25" (203mm) 6.75" (111mm)
7.5
20 2H
Curves based on clear water with specific gravity of 1.0
P(1
0 0
100
15H
P(1
55% 50%
9.75" (216mm)
65% 60%
74% 72% 70%
60
200
300
5H
.5k
3H W)
P(3
P(2
.2k
10H
HP
.7k
W)
400 500 600 Flow in gallons per minute
P(7
(6.
6kW
W)
700
)
1.2
kW
)
10
.5k
W)
800
5
900
0 1000
100
Head in kilopascals
78% 76%
11.25" (229mm)
Head in meters
76%
78%
80%
Head in feet
80
70% 72% 74%
50% 55%
60% 65%
12.75* (241.3mm) 12.75"
50
0
FIGURE 14–15 Example of a manufacturer’s performance plot for a family of centrifugal pumps. Each pump has the same casing, but a different impeller diameter. Courtesy of Taco, Inc., Cranston, RI. Used by permission.
EXAMPLE 14–2
Selection of Pump Impeller Size
A washing operation at a power plant requires 370 gallons per minute (gpm) of water. The required net head is about 24 ft at this flow rate. A newly hired engineer looks through some catalogs and decides to purchase the 8.25-in impeller option of the Taco Model 4013 FI Series centrifugal pump of Fig. 14–15. If the pump operates at 1160 rpm, as specified in the performance plot, she reasons, its performance curve intersects 370 gpm at H ! 24 ft. The chief engineer, who is very concerned about efficiency, glances at the performance curves and notes that the efficiency of this pump at this operating point is only 70 percent. He sees that the 12.75-in impeller option achieves a higher efficiency (about 76.5 percent) at the same flow rate. He notes that a throttle valve can be installed downstream of the pump to increase the required net head so that the pump operates at this higher efficiency. He asks the junior engineer to justify her choice of impeller diameter. Namely, he asks her to calculate which impeller option
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745 CHAPTER 14
(8.25-in or 12.75-in) would need the least amount of electricity to operate (Fig. 14–16). Perform the comparison and discuss.
SOLUTION For a given flow rate and net head, we are to calculate which impeller size uses the least amount of power, and we are to discuss our results. Assumptions 1 The water is at 70°F. 2 The flow requirements (volume flow rate and head) are constant. Properties For water at 70°F, r ! 62.30 lbm/ft3. Analysis From the contours of brake horsepower that are shown on the performance plot of Fig. 14–15, the junior engineer estimates that the pump with the smaller impeller requires about 3.2 hp from the motor. She verifies this estimate by using Eq. 14–5,
Is she trying to tell me that the less efficient pump can actually save on energy costs?
Required bhp for the 8.25-in impeller option: # rgV H (62.30 lbm/ft3)(32.2 ft/s2)(370 gal/min)(24 ft) bhp ! ! h pump 0.70 hp ' s 0.1337 ft3 lbf 1 min ba ba ba b ! 3.20 hp 2 gal 60 s 550 ft ' lbf 32.2 lbm ' ft/s
%a
Similarly, the larger-diameter impeller option requires
Required bhp for the 12.75-in impeller option:
bhp ! 8.78 hp
. using the operating point of that pump, namely, V ! 370 gpm, H ! 72.0 ft, and hpump ! 76.5 percent (Fig. 14–15). Clearly, the smaller-diameter impeller option is the better choice in spite of its lower efficiency, because it uses less than half the power. Discussion Although the larger impeller pump would operate at a somewhat higher value of efficiency, it would deliver about 72 ft of net head at the required flow rate. This is overkill, and the throttle valve would be required to make up the difference between this net head and the required flow head of 24 ft of water. A throttle valve does nothing more than waste mechanical energy, however; so the gain in efficiency of the pump is more than offset by losses through the throttle valve. If the flow head or capacity requirements increase at some time in the future, a larger impeller can be purchased for the same casing.
Pump Cavitation and Net Positive Suction Head
When pumping liquids, it is possible for the local pressure inside the pump to fall below the vapor pressure of the liquid, Pv. (Pv is also called the saturation pressure Psat and is listed in thermodynamics tables as a function of saturation temperature.) When P & Pv, vapor-filled bubbles called cavitation bubbles appear. In other words, the liquid boils locally, typically on the suction side of the rotating impeller blades where the pressure is lowest (Fig. 14–17). After the cavitation bubbles are formed, they are transported through the pump to regions where the pressure is higher, causing rapid collapse of the bubbles. It is this collapse of the bubbles that is undesirable, since it causes noise, vibration, reduced efficiency, and most importantly, damage to the impeller blades. Repeated bubble collapse near a blade surface leads to pitting or erosion of the blade and eventually catastrophic blade failure.
FIGURE 14–16 In some applications, a less efficient pump from the same family of pumps may require less energy to operate. An even better choice, however, would be a pump whose best efficiency point occurs at the required operating point of the pump, but such a pump is not always commercially available.
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746 FLUID MECHANICS Cavitation bubbles collapse
Cavitation bubbles form
Pressure side Impeller blade Suction side
Net positive suction head: v
FIGURE 14–17 Cavitation bubbles forming and collapsing on the suction side of an impeller blade.
Head
H
NPSH required 0
•
V
0
FIGURE 14–18 Typical pump performance curve in which net head and required net positive suction head are plotted versus volume flow rate.
H
Head
No cavitation
Cavitation
NPSH NPSH required
0 0
To avoid cavitation, we must ensure that the local pressure everywhere inside the pump stays above the vapor pressure. Since pressure is most easily measured (or estimated) at the inlet of the pump, cavitation criteria are typically specified at the pump inlet. It is useful to employ a flow parameter called net positive suction head (NPSH), defined as the difference between the pump’s inlet stagnation pressure head and the vapor pressure head,
•
Vmax
NPSH ! a
Pv V2 P $ b # rg 2g pump inlet rg
(14–8)
Pump manufacturers test their pumps for cavitation in a pump test facility by varying the volume flow rate and inlet pressure in a controlled manner. Specifically, at a given flow rate and liquid temperature, the pressure at the pump inlet is slowly lowered until cavitation occurs somewhere inside the pump. The value of NPSH is calculated using Eq. 14–8 and is recorded at this operating condition. The process is repeated at several other flow rates, and the pump manufacturer then publishes a performance parameter called the required net positive suction head (NPSHrequired), defined as the minimum NPSH necessary to avoid cavitation in the pump. The measured value of NPSHrequired varies with volume flow rate, and therefore NPSHrequired is often plotted on the same pump performance curve as net head (Fig. 14–18). When expressed properly in units of head of the liquid being pumped, NPSHrequired is independent of the type of liquid. However, if the required net positive suction head is expressed for a particular liquid in pressure units such as pascals or psi, the engineer must be careful to convert this pressure to the equivalent column height of the actual liquid being pumped. Note that since NPSHrequired is usually much smaller than H over the majority of the performance curve, it is often plotted on a separate expanded vertical axis for clarity (see Fig. 14–15) or as contour lines when being shown for a family of with volume flow rate, although for pumps. NPSHrequired typically increases . some pumps it decreases with V at low flow rates where the pump is not operating very efficiently, as sketched in Fig. 14–18. In order to ensure that a pump does not cavitate, the actual or available NPSH must be greater than NPSHrequired. It is important to note that the value of NPSH varies not only with flow rate, but also with liquid temperature, since Pv is a function of temperature. NPSH also depends on the type of liquid being pumped, since there is a unique Pv versus T curve for each liquid. Since irreversible head losses through the piping system upstream of the inlet increase with flow rate, the pump inlet stagnation pressure head . decreases with flow rate. Therefore, the value of NPSH decreases with V, as sketched in Fig. 14–19. By identifying the volume flow rate at which the curves of actual NPSH and NPSHrequired intersect, we estimate the maximum volume flow rate that can be delivered by the pump without cavitation (Fig. 14–19).
•
V
FIGURE 14–19 The volume flow rate at which the actual NPSH and the required NPSH intersect represents the maximum flow rate that can be delivered by the pump without the occurrence of cavitation.
EXAMPLE 14–3
Maximum Flow Rate to Avoid Pump Cavitation
The 11.25-in impeller option of the Taco Model 4013 FI Series centrifugal pump of Fig. 14–15 is used to pump water at 25°C from a reservoir whose surface is 4.0 ft above the centerline of the pump inlet (Fig. 14–20). The piping system from the reservoir to the pump consists of 10.5 ft of cast iron pipe with an ID of 4.0 in and an average inner roughness height of 0.02 in.
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747 CHAPTER 14
There are several minor losses: a sharp-edged inlet (KL ! 0.5), three flanged smooth 90° regular elbows (KL ! 0.3 each), and a fully open flanged globe valve (KL ! 6.0). Estimate the maximum volume flow rate (in units of gpm) that can be pumped without cavitation. If the water were warmer, would this maximum flow rate increase or decrease? Why? Discuss how you might increase the maximum flow rate while still avoiding cavitation.
SOLUTION For a given pump and piping system we are to estimate the maximum volume flow rate that can be pumped without cavitation. We are also to discuss the effect of water temperature and how we might increase the maximum flow rate. Assumptions 1 The flow is steady. 2 The liquid is incompressible. 3 The flow at the pump inlet is turbulent and fully developed, with a ! 1.05. Properties For water at T ! 25°C, r ! 997.0 kg/m3, m ! 8.91 % 10#4 kg/m · s, and Pv ! 3.169 kPa. Standard atmospheric pressure is Patm ! 101.3 kPa. Analysis We apply the steady energy equation in head form along a streamline from point 1 at the reservoir surface to point 2 at the pump inlet, P1 a 1V 21 P2 a 2V 22 $ $ $ z 1 $ hpump, u ! $ z 2 $ hturbine, e $ hL, total rg rg 2g 2g
(1)
In Eq. 1 we have ignored the water speed at the reservoir surface (V1 ≅ 0). There is no turbine in the piping system. Also, although there is a pump in the system, there is no pump between points 1 and 2; hence the pump head term also drops out. We solve Eq. 1 for P2 /rg, which is the pump inlet pressure expressed as a head,
Pump inlet pressure head:
P2 Patm a 2V 22 ! $ (z 1 # z 2) # # hL, total rg rg 2g
(2)
Note that in Eq. 2, we have recognized that P1 ! Patm since the reservoir surface is exposed to atmospheric pressure. The available net positive suction head at the pump inlet is obtained from Eq. 14–8. After substitution of Eq. 2, we get
Available NPSH: NPSH !
Patm # Pv (a 2 # 1)V 22 $ (z 1 # z 2) # hL, total # (3) rg 2g
Since we know Patm, Pv , and the elevation difference, all that remains is to estimate the total irreversible head loss through the piping system, which depends on volume flow rate. Since the pipe diameter is constant,
Irreversible head loss:
hL, total ! af
L V2 $ a K Lb D 2g
(4)
The rest of the problem is most easily solved on a computer. For a given volume flow rate, we calculate speed V and Reynolds number Re. From Re and the known pipe roughness, we use the Moody chart (or the Colebrook equation) to obtain friction factor f. The sum of all the minor loss coefficients is
Minor losses:
a K L ! 0.5 $ 3 % 0.3 $ 6.0 ! 7.4
(5)
. We make one calculation by hand for illustrative purposes. At V ! 400 gpm (0.02523 m3/s), the average speed of water through the pipe is
# # 2 4(0.02523 m3/s) V 4V 1 in V! ! ! a b ! 3.112 m/s A pD2 0.0254 m p(4.0 in)2
(6)
1 z1 Reservoir
Inlet piping system
Pump z2 Valve
2
FIGURE 14–20 Inlet piping system from the reservoir (1) to the pump inlet (2) for Example 14–3.
cen72367_ch14.qxd 11/12/04 2:11 PM Page 748
748 FLUID MECHANICS 30
25
Available NPSH, 60°C
20 NPSH, ft
which produces a Reynolds number of Re ! rVD/m ! 3.538 % 105. At this Reynolds number, and with roughness factor e/D ! 0.005, the Colebrook equation yields f ! 0.0306. Substituting the given properties, along with f, D, L, and Eqs. 4, 5, and 6, into Eq. 3, we calculate the available net positive suction head at this flow rate,
Available NPSH, 25°C
No cavitation, T = 25°C
15
Required NPSH
10
NPSH !
(101,300 # 3169) N/m2 kg ' m/s2 b $ 1.219 m a N (997.0 kg/m3)(9.81 m/s2) # a0.0306
(3.112 m/s)2 10.5 ft $ 7.4 # (1.05 # 1)b 0.3333 ft 2(9.81 m/s2)
! 7.148 m ! 23.5 ft 5 No cavitation, T = 60°C
0 300
400
500 • V, gpm
600
700
FIGURE 14–21 Net positive suction head as a function of volume flow rate for the pump of Example 14–3 at two temperatures. Cavitation is predicted to occur at flow rates greater than the point where the available and required values of NPSH intersect.
(a)
The required net positive suction head is obtained from Fig. 14–15. At our example flow rate of 400 gpm, NPSHrequired is just above 4.0 ft. Since the actual NPSH is much higher than this, we need not worry about cavitation at this flow rate. We use EES (or a spreadsheet) to calculate NPSH as a function of volume flow rate, and the results are plotted in Fig. 14–21. It is clear from this plot that at 25°C, cavitation occurs at flow rates above approximately 600 gpm—close to the free delivery. If the water were warmer than 25°C, the vapor pressure would increase, the viscosity would decrease, and the density would decrease slightly. The calculations are repeated at T ! 60°C, at which r ! 983.3 kg/m3, m ! 4.67 % 10#4 kg/m · s, and Pv ! 19.94 kPa. The results are also plotted in Fig. 14–21, where we see that the maximum volume flow rate without cavitation decreases with temperature (to about 555 gpm at 60°C). This decrease agrees with our intuition, since warmer water is already closer to its boiling point from the start. Finally, how can we increase the maximum flow rate? Any modification that increases the available NPSH helps. We can raise the height of the reservoir surface (to increase the hydrostatic head). We can reroute the piping so that only one elbow is necessary and replace the globe valve with a ball valve (to decrease the minor losses). We can increase the diameter of the pipe and decrease the surface roughness (to decrease the major losses). In this particular problem, the minor losses have the greatest influence, but in many problems, the major losses are more significant, and increasing the pipe diameter is most effective. That is one reason why many centrifugal pumps have a larger inlet diameter than outlet diameter. Discussion Note that NPSHrequired does not depend on water temperature, but the actual or available NPSH decreases with temperature (Fig. 14–21).
Pumps in Series and Parallel
(b)
FIGURE 14–22 Arranging two very dissimilar pumps in (a) series or (b) parallel can sometimes lead to problems.
(7)
When faced with the need to increase volume flow rate or pressure rise by a small amount, you might consider adding an additional smaller pump in series or in parallel with the original pump. While series or parallel arrangement is acceptable for some applications, arranging dissimilar pumps in series or in parallel may lead to problems, especially if one pump is much larger than the other (Fig. 14–22). A better course of action is to increase the original pump’s speed and/or input power (larger electric motor), replace the impeller with a larger one, or replace the entire pump with a larger one. The logic for this decision can be seen from the pump performance curves, realizing that pressure rise and volume flow rate are related. Arranging dis-
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749 CHAPTER 14
similar pumps in series may create problems because the volume flow rate through each pump must be the same, but the overall pressure rise is equal to the pressure rise of one pump plus that of the other. If the pumps have widely different performance curves, the smaller pump may be forced to operate beyond its free delivery flow rate, whereupon it acts like a head loss, reducing the total volume flow rate. Arranging dissimilar pumps in parallel may create problems because the overall pressure rise must be the same, but the net volume flow rate is the sum of that through each branch. If the pumps are not sized properly, the smaller pump may not be able to handle the large head imposed on it, and the flow in its branch could actually be reversed; this would inadvertently reduce the overall pressure rise. In either case, the power supplied to the smaller pump would be wasted. Keeping these cautions in mind, there are many applications where two or more similar (usually identical) pumps are operated in series or in parallel. When operated in series, the combined net head is simply the sum of the net heads of each pump (at a given volume flow rate), Hcombined ! a Hi n
Combined net head for n pumps in series:
(14–9)
i!1
Equation 14–9 is illustrated in Fig. 14–23 for three pumps in series. In this example, pump 3 is the strongest and pump 1 is the weakest. The shutoff head of the three pumps combined in series is equal to the sum of the shutoff head of each individual pump. For low values of volume flow rate, the net head of the three pumps in series is equal to H1 $ H2 $ H3. Beyond the free delivery of pump 1 (to the right of the first vertical dashed gray line in Fig. 14–23), pump 1 should be shut off and bypassed. Otherwise it would be running beyond its maximum designed operating point, and the pump or its motor could be damaged. Furthermore, the net head across this pump would be negative as previously discussed, contributing to a net loss in the system. With pump 1 bypassed, the combined net head becomes H2 $ H3. Similarly, beyond the free delivery of pump 2, that pump should also be shut off and bypassed, and the combined net head is then equal to H3 alone, as indicated to the right of the second vertical dashed gray line in Fig. 14–23. In this
Shutoff head of combined pumps H
Pump 1 should be shut off and bypassed
H1 + H2 + H3 Combined net head
H2 + H3
Pump 2 should be shut off and bypassed H3 only
Pump 1 Pump 2 0 0
Pump 3 •
Free delivery of combined pumps
V
FIGURE 14–23 Pump performance curve (dark blue) for three dissimilar pumps in series. At low values of volume flow rate, the combined net head is equal to the sum of the net head of each pump by itself. However, to avoid pump damage and loss of combined net head, any individual pump should be shut off and bypassed at flow rates larger than that pump’s free delivery, as indicated by the vertical dashed gray lines. If the three pumps were identical, it would not be necessary to turn off any of the pumps, since the free delivery of each pump would occur at the same volume flow rate.
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750 FLUID MECHANICS
case, the combined free delivery is the same as that of pump 3 alone, assuming that the other two pumps are bypassed. When two or more identical (or similar) pumps are operated in parallel, their individual volume flow rates (rather than net heads) are summed, n # # Combined capacity for n pumps in parallel: V combined ! a V i
(14–10)
i!1
As an example, consider the same three pumps, but arranged in parallel rather than in series. The combined pump performance curve is shown in Fig. 14–24. The free delivery of the three combined pumps is equal to the sum of the free delivery of each individual pump. For low .values. of net . head, the capacity of the three pumps in parallel is equal to V1 $ V2 $ V3. Above the shutoff head of pump 1 (above the first horizontal dashed gray line in Fig. 14–24), pump 1 should be shut off and its branch should be blocked (with a valve). Otherwise it would be running beyond its maximum designed operating point, and the pump or its motor could be damaged. Furthermore, the volume flow rate through this pump would be negative as previously discussed, contributing to a net loss in the system. With pump 1 . . shut off and blocked, the combined capacity becomes V2 $ V3. Similarly, above the shutoff head of pump 2, that pump should. also be shut off and blocked. The combined capacity is then equal to V3 alone, as indicated above, the second horizontal dashed gray line in Fig. 14–24. In this case, the combined shutoff head is the same as that of pump 3 alone, assuming that the other two pumps are shut off and their branches are blocked. In practice, several pumps may be combined in parallel to deliver a large volume flow rate (Fig. 14–25). Examples include banks of pumps used to circulate water in cooling towers and chilled water loops (Wright, 1999). Ideally all the pumps should be identical so that we don’t need to worry
FIGURE 14–24 Pump performance curve (dark blue) for three pumps in parallel. At a low value of net head, the combined capacity is equal to the sum of the capacity of each pump by itself. However, to avoid pump damage and loss of combined capacity, any individual pump should be shut off at net heads larger than that pump’s shutoff head, as indicated by the horizontal dashed gray lines. That pump’s branch should also be blocked with a valve to avoid reverse flow. If the three pumps were identical, it would not be necessary to turn off any of the pumps, since the shutoff head of each pump would occur at the same net head.
Shutoff head of combined pumps H ⋅ V3 only
Pump 2 should be shut off ⋅ ⋅ V2 + V3
Pump 1 should be shut off ⋅ ⋅ ⋅ V1 + V2 + V3
Pump 2 Pump 1
Pump 3 Combined capacity
0 0
Free delivery of combined pumps
⋅ V
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751 CHAPTER 14
FIGURE 14–25 Several identical pumps are often run in a parallel configuration so that a large volume flow rate can be achieved when necessary. Three parallel pumps are shown. Courtesy of Goulds Pumps, ITT Industries. Used by permission.
about shutting any of them off (Fig. 14–24). Also, it is wise to install check valves in each branch so that when a pump needs to be shut down (for maintenance or when the required flow rate is low), backflow through the pump is avoided. Note that the extra valves and piping required for a parallel pump network add additional head losses to the system; thus the overall performance of the combined pumps suffers somewhat.
Positive-Displacement Pumps
People have designed numerous positive-displacement pumps throughout the centuries. In each design, fluid is sucked into an expanding volume and then pushed along as that volume contracts, but the mechanism that causes this change in volume differs greatly among the various designs. Some designs are very simple, like the flexible-tube peristaltic pump (Fig. 14–26a) that compresses a tube by small wheels, pushing the fluid along. (This mechanism is somewhat similar to peristalsis in your esophagus or intestines, where muscles rather than wheels compress the tube.) Others are more complex, using rotating cams with synchronized lobes (Fig. 14–26b), interlocking gears (Fig. 14–26c), or screws (Fig. 14–26d ). Positive-displacement pumps are ideal for high-pressure applications like pumping viscous liquids or thick slurries, and for applications where precise amounts of liquid are to be dispensed or metered, as in medical applications.
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752 FLUID MECHANICS
(a)
(b)
(c)
(d)
FIGURE 14–26 Examples of positive-displacement pumps: (a) flexible-tube peristaltic pump, (b) three-lobe rotary pump, (c) gear pump, and (d) double screw pump. Adapted from F. M. White, Fluid Mechanics 4/e. Copyright © 1999. The McGraw-Hill Companies, Inc. With permission.
FIGURE 14–27 Four phases (one-eighth of a turn apart) in the operation of a two-lobe rotary pump, a type of positivedisplacement pump. The light blue region represents a chunk of fluid pushed through the top rotor, while the dark blue region represents a chunk of fluid pushed through the bottom rotor, which rotates in the opposite direction. Flow is from left to right.
In
Out
45°
90°
135°
180°
To illustrate the operation of a positive-displacement pump, we sketch four phases of half of a cycle of a simple rotary pump with two lobes on each rotor (Fig. 14–27). The two rotors are synchronized by an external gear box so as to rotate at the same angular speed, but in opposite directions. In the diagram, the top rotor turns clockwise and the bottom rotor turns counterclockwise, sucking in fluid from the left and discharging it to the right. A white dot is drawn on one lobe of each rotor to help you visualize the rotation.
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Gaps exist between the rotors and the housing and between the lobes of the rotors themselves, as illustrated (and exaggerated) in Fig. 14–27. Fluid can leak through these gaps, reducing the pump’s efficiency. High-viscosity fluids cannot penetrate the gaps as easily; hence the net head (and efficiency) of a rotary pump generally increases with fluid viscosity, as shown in Fig. 14–28. This is one reason why rotary pumps (and other types of positivedisplacement pumps) are a good choice for pumping highly viscous fluids and slurries. They are used, for example, as automobile engine oil pumps and in the foods industry to pump heavy liquids like syrup, tomato paste, and chocolate, and slurries like soups. The pump performance curve (net head versus capacity) of a rotary pump is nearly vertical throughout its recommended operating range, since the capacity is fairly constant regardless of load at a given rotational speed (Fig. 14–28). However, as indicated by the dashed blue line in Fig. 14–28, at very high values of net head, corresponding to very high pump outlet pressure, leaks become more severe, even for high-viscosity fluids. In addition, the motor driving the pump cannot overcome the large torque caused by this high outlet pressure, and the motor begins to suffer stall or overload, which may burn out the motor. Therefore, rotary pump manufacturers do not recommend operation of the pump above a certain maximum net head, which is typically well below the shutoff head. The pump performance curves supplied by the manufacturer often do not even show the pump’s performance outside of its recommended operating range. Positive-displacement pumps have many advantages over dynamic pumps. For example, a positive-displacement pump is better able to handle shear sensitive liquids since the induced shear is much less than that of a dynamic pump operating at similar pressure and flow rate. Blood is a shear sensitive liquid, and this is one reason why positive-displacement pumps are used for artificial hearts. A well-sealed positive-displacement pump can create a significant vacuum pressure at its inlet, even when dry, and is thus able to lift a liquid from several meters below the pump. We refer to this kind of pump as a self-priming pump (Fig. 14–29). Finally, the rotor(s) of a positivedisplacement pump run at lower speeds than the rotor (impeller) of a dynamic pump at similar loads, extending the useful lifetime of seals, etc. There are some disadvantages of positive-displacement pumps as well. Their volume flow rate cannot be changed unless the rotation rate is changed. (This is not as simple as it sounds, since most AC electric motors are designed to operate at one or more fixed rotational speeds.) They create very high pressure at the outlet side, and if the outlet becomes blocked, ruptures may occur or electric motors may overheat, as previously discussed. Overpressure protection (e.g., a pressure-relief valve) is often required for this reason. Because of their design, positive-displacement pumps may deliver a pulsating flow, which may be unacceptable for some applications. Analysis of positive-displacement pumps is fairly straightforward. From the geometry of the pump, we calculate the closed volume (Vclosed) that is filled (and expelled) for every n rotations of the shaft. Volume flow rate is . then equal to rotation rate n times Vclosed divided by n, Volume flow rate, positive-displacement pump:
# # Vclosed V !n n
(14–11)
Shutoff head Maximum recommended net head
H
Recommended operating range Increasing velocity
0
•
0
Free delivery
V
FIGURE 14–28 Comparison of the pump performance curves of a rotary pump operating at the same speed, but with fluids of various viscosities. To avoid motor overload the pump should not be operated in the shaded region.
Self-priming pump
Out
Hose In
FIGURE 14–29 A pump that can lift a liquid even when the pump itself is “empty” is called a self-priming pump.
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EXAMPLE 14–4 ⋅ V
In
⋅ V
Out
V lobe
FIGURE 14–30 The two-lobe rotary pump of Example 14–4. Flow is from left to right.
Blade Impeller shroud Flow in v
(a)
SOLUTION We are to calculate the volume flow rate of oil through a positivedisplacement pump for given values of lobe volume and rotation rate. Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between lobes or between lobes and the casing. 3 The oil is incompressible. Analysis By studying Fig. 14–27, we see that for half of a rotation (180° for n ! 0.5 rotations) of the two counter-rotating shafts, the total volume of oil pumped is Vclosed ! 2Vlobe. The volume flow rate is then calculated from Eq. 14–11,
Discussion If there were leaks in the pump, the volume flow rate would be lower. The oil’s density is not needed for calculation of the volume flow rate. However, the higher the fluid density, the higher the required shaft torque and brake horsepower.
Dynamic Pumps
Flow out Impeller shroud
Flow in v (b) Blade
A two-lobe rotary positive-displacement pump, similar to that of Fig. 14–27, moves 0.45 cm3 of SAE 30 motor oil in each lobe volume Vlobe, as sketched . in Fig. 14–30. Calculate the volume flow rate of oil for the case where n ! 900 rpm.
# 2(0.45 cm3) # Vclosed ! (900 rot/min) V !n ! 1620 cm3/min n 0.5 rot
Flow out
Blade
Volume Flow Rate through a Positive-Displacement Pump
Flow out
Flow in Impeller hub
v
(c)
FIGURE 14–31 The impeller (rotating portion) of the three main categories of dynamic pumps: (a) centrifugal flow, (b) mixed flow, and (c) axial flow.
There are three main types of dynamic pumps that involve rotating blades called impeller blades or rotor blades, which impart momentum to the fluid. For this reason they are sometimes called rotodynamic pumps or simply rotary pumps (not to be confused with rotary positive-displacement pumps, which use the same name). There are also some nonrotary dynamic pumps, such as jet pumps and electromagnetic pumps; these are not discussed in this text. Rotary pumps are classified by the manner in which flow exits the pump: centrifugal flow, axial flow, and mixed flow (Fig. 14–31). In a centrifugal-flow pump, fluid enters axially (in the same direction as the axis of the rotating shaft) in the center of the pump, but is discharged radially (or tangentially) along the outer radius of the pump casing. For this reason centrifugal pumps are also called radial-flow pumps. In an axial-flow pump, fluid enters and leaves axially, typically along the outer portion of the pump because of blockage by the shaft, motor, hub, etc. A mixed-flow pump is intermediate between centrifugal and axial, with the flow entering axially, not necessarily in the center, but leaving at some angle between radially and axially.
Centrifugal Pumps
Centrifugal pumps and blowers can be easily identified by their snail-shaped casing, called the scroll (Fig. 14–32). They are found all around your home—in dishwashers, hot tubs, clothes washers and dryers, hairdryers,
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vacuum cleaners, kitchen exhaust hoods, bathroom exhaust fans, leaf blowers, furnaces, etc. They are used in cars—the water pump in the engine, the air blower in the heater/air conditioner unit, etc. Centrifugal pumps are ubiquitous in industry as well; they are used in building ventilation systems, washing operations, cooling ponds and cooling towers, and in numerous other industrial operations in which fluids are pumped. A schematic diagram of a centrifugal pump is shown in Fig. 14–33. Note that a shroud often surrounds the impeller blades to increase blade stiffness. In pump terminology, the rotating assembly that consists of the shaft, the hub, the impeller blades, and the impeller shroud is called the impeller or rotor. Fluid enters axially through the hollow middle portion of the pump (the eye), after which it encounters the rotating blades. It acquires tangential and radial velocity by momentum transfer with the impeller blades, and acquires additional radial velocity by so-called centrifugal forces, which are actually a lack of sufficient centripetal forces to sustain circular motion. The flow leaves the impeller after gaining both speed and pressure as it is flung radially outward into the scroll (also called the volute). As sketched in Fig. 14–33, the scroll is a snail-shaped diffuser whose purpose is to decelerate the fast-moving fluid leaving the trailing edges of the impeller blades, thereby further increasing the fluid’s pressure, and to combine and direct the flow from all the blade passages toward a common outlet. As mentioned previously, if the flow is steady in the mean, if the fluid is incompressible, and if the inlet and outlet diameters are the same, the average flow speed at the outlet is identical to that at the inlet. Thus, it is not necessarily the speed, but the pressure that increases from inlet to outlet through a centrifugal pump. There are three types of centrifugal pump that warrant discussion, based on impeller blade geometry, as sketched in Fig. 14–34: backward-inclined blades, radial blades, and forward-inclined blades. Centrifugal pumps with backward-inclined blades (Fig. 14-34a) are the most common. These yield the highest efficiency of the three because fluid flows into and out of the blade passages with the least amount of turning. Sometimes the blades are airfoil shaped, yielding similar performance but even higher efficiency. The pressure rise is intermediate between the other two types of centrifugal
Casing
Impeller shroud
b2
Vout, Pout
r2
b1 r1
In Pin
Out
v
Vin
v
In
Shaft Impeller blade
Impeller
Eye
Scroll Side view
Frontal view
FIGURE 14–32 A typical centrifugal blower with its characteristic snail-shaped scroll. Courtesy The New York Blower Company, Willowbrook, IL. Used by permission.
FIGURE 14–33 Side view and frontal view of a typical centrifugal pump. Fluid enters axially in the middle of the pump (the eye), is flung around to the outside by the rotating blade assembly (impeller), is diffused in the expanding diffuser (scroll), and is discharged out the side of the pump. We define r1 and r2 as the radial locations of the impeller blade inlet and outlet, respectively; b1 and b2 are the axial blade widths at the impeller blade inlet and outlet, respectively.
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756 FLUID MECHANICS
v
(a)
v
(b)
(c) Backward Forward Radial H H or bhp
bhp 0
V
0 (d)
FIGURE 14–34 The three main types of centrifugal pumps are those with (a) backwardinclined blades, (b) radial blades, and (c) forward-inclined blades; (d) comparison of net head and brake horsepower performance curves for the three types of centrifugal pumps.
pumps. Centrifugal pumps with radial blades (also called straight blades, Fig. 14–34b) have the simplest geometry and produce the largest pressure rise of the three for a wide range of volume flow rates, but the pressure rise decreases rapidly after the point of maximum efficiency. Centrifugal pumps with forward-inclined blades (Fig. 14–34c) produce a pressure rise that is nearly constant, albeit lower than that of radial or backward-inclined blades, over a wide range of volume flow rates. Forward-inclined centrifugal pumps generally have more blades, but the blades are smaller, as sketched in Fig. 14–34c. Centrifugal pumps with forward-inclined blades generally have a lower maximum efficiency than do straight-bladed pumps. Radial and backward-inclined centrifugal pumps are preferred for applications where one needs to provide volume flow rate and pressure rise within a narrow range of values. If a wider range of volume flow rates and/or pressure rises are desired, the performance of radial pumps and backward-inclined pumps may not be able to satisfy the new requirements; these types of pumps are less forgiving (less robust). The performance of forward-inclined pumps is more forgiving and accommodates a wider variation, at the cost of lower efficiency and less pressure rise per unit of input power. If a pump is needed to produce large pressure rise over a wide range of volume flow rates, the forward-inclined centrifugal pump is attractive. Net head and brake horsepower performance curves for these three types of centrifugal pump are compared in Fig. 14–34d. The curves have been adjusted such that each pump achieves the same free delivery (maximum volume flow rate at zero net head). Note that these are qualitative sketches for comparison purposes only—actual measured performance curves may differ significantly in shape, depending on details of the pump design. For any inclination of the impeller blades (backward, radial, or forward), we can analyze the velocity vectors through the blades. The actual flow field is unsteady, fully three-dimensional, and perhaps compressible. For simplicity in our analysis we consider steady flow in both the absolute reference frame and in the relative frame of reference rotating with the impeller. We consider only incompressible flow, and we consider only the radial or normal velocity component (subscript n) and the circumferential or tangential velocity component (subscript t) from blade inlet to blade outlet. We do not consider the axial velocity component (to the right in Fig. 14–35 and into the page in the frontal view of Fig. 14–33). In other words, although there is a nonzero axial component of velocity through the impeller, it does not enter our analysis. A close-up side view of a simplified centrifugal pump is sketched in Fig. 14–35, where we define V1, n and V2, n as the average normal components of velocity at radii r1 and r2, respectively. Although a gap is shown between the blade and the casing, we assume in our simplified analysis that . no leakage occurs in these gaps. The volume flow rate V entering the eye of the pump passes through the circumferential cross-sectional area defined by width b1 at radius r1. Conservation of mass requires that this same volume flow rate must pass through the circumferential cross-sectional area defined by width b2 at radius r2. Using the average normal velocity components V1, n and V2, n defined in Fig. 14–35, we write Volume flow rate:
# V ! 2pr1b1V1, n ! 2pr2b2V2, n
(14–12)
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757 CHAPTER 14
from which we obtain V2, n ! V1, n
r1b1 r2b2
Scroll (14–13)
It is clear from Eq. 14–13 that V2, n may be less than, equal to, or greater than V1, n, depending on the values of b and r at the two radii. We sketch a close-up frontal view of one impeller blade in Fig. 14–36, where we show both radial and tangential velocity components. We have drawn a backward-inclined blade, but the same analysis holds for blades of any inclination. The inlet of the blade (at radius r1) moves at tangential velocity vr1. Likewise, the outlet of the blade moves at tangential velocity vr2. It is clear from Fig. 14–36 that these two tangential velocities differ not only in magnitude, but also in direction, because of the inclination of the blade. We define leading edge angle b1 as the blade angle relative to the reverse tangential direction at radius r1. In like manner we define trailing edge angle b2 as the blade angle relative to the reverse tangential direction at radius r2. We now make a significant simplifying approximation. We assume that the flow impinges on the blade parallel to the blade’s leading edge and exits the blade parallel to the blade’s trailing edge. In other words,
Impeller blade
b2 r2 b1 r1
V1, n
•
Vin, Pin, V In
v
Shaft
FIGURE 14–35 Close-up side view of the simplified centrifugal flow pump used for elementary analysis of the velocity vectors; V1, n and V2, n are defined as the average normal (radial) components of velocity at radii r1 and r2, respectively.
We assume that the flow is everywhere tangent to the blade surface when viewed from a reference frame rotating with the blade.
At the inlet, this approximation is sometimes called the shockless entry condition, not to be confused with shock waves (Chap. 12). Rather, the terminology implies smooth flow into the impeller blade without a sudden turning “shock.” Inherent in this approximation is the assumption that there is no flow separation anywhere along the blade surface. If the centrifugal pump operates at or near its design conditions, this assumption is valid. However, when the pump operates far off design conditions, the flow may separate off the blade surface (typically on the suction side where there are adverse pressure gradients), and →our simplified analysis breaks down. → Velocity vectors V 1, relative and V 2, relative are drawn in Fig. 14–36 parallel to the blade surface, in accordance with our simplifying assumption. These are the velocity vectors seen from the relative reference frame of an observer moving with the rotating blade. When we vectorially add tangential velocity → vr1 (the velocity of the blade at radius r1) to V 1, relative by completing the parallelogram as sketched in Fig. 14–36, the resultant vector is the absolute → → fluid velocity V 1 at the blade inlet. In exactly similar fashion, we obtain V 2, the absolute fluid velocity at the blade outlet (also sketched in Fig. 14–36). For completeness, normal velocity components V1, n and V2, n are also shown in Fig. 14–36. Notice that these normal velocity components are independent of which frame of reference we use, absolute or relative. To evaluate the torque on the rotating shaft, we apply the angular momentum relation for a control volume, as discussed in Chap. 6. We choose a control volume surrounding the impeller blades, from radius r1 to radius r2, as sketched in Fig. 14–37. We also introduce in Fig. 14–37 angles a1 and a2, defined as the angle of departure of the absolute velocity vector from the normal direction at radii r1 and r2, respectively. In keeping with the concept of treating a control volume like a “black box,” we ignore details of individual impeller blades. Instead we assume that flow enters the control volume
Casing
V2, n Out
V2, t
→
V2, relative
V2
V2, n b2
vr2
b2 →
→
b2
V1, relative
V2, relative
V1, n → V1
b1 r2
b1 v
r1
vr1
FIGURE 14–36 Close-up frontal view of the simplified centrifugal flow pump used for elementary analysis of the velocity vectors. Absolute velocity vectors of the fluid are shown as bold arrows. It is assumed that the flow is everywhere tangent to the blade surface when viewed from a reference frame rotating with the blade, as indicated by the relative velocity vectors.
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758 FLUID MECHANICS →
→
V2, n
V2
a2
V2, t
r2
a1
V1, n r1
Control volume
→
V1 V1, t
Shaft
v
Tshaft = torque supplied to shaft
FIGURE 14–37 Control volume (shaded) used for angular momentum analysis of a centrifugal pump; absolute tangential velocity components V1, t and V2, t are labeled.
with uniform absolute velocity V 1 around the→ entire circumference at radius r1 and exits with uniform absolute velocity V 2 around the entire circumference at radius r2. → → Since moment of momentum →is defined as the cross product r % V , only → the tangential components of V 1 and V 2 are relevant to the shaft torque. These are shown as V1, t and V2, t in Fig. 14–37. It turns out that shaft torque is equal to the change in moment of momentum from inlet to outlet, as given by the Euler turbomachine equation (also called Euler’s turbine formula), derived in Chap. 6, Euler turbomachine equation:
Alternative form, Euler turbomachine equation: # Tshaft ! rV (r2V2 sin a 2 # r1V1 sin a 1)
V2
a2
V2, t →
V1
r2 r1 v
Control volume
FIGURE 14–38 Control volume and absolute velocity vectors for the centrifugal blower of Example 14–5. The view is along the blower axis.
(14-15)
In our simplified analysis there are no irreversible losses. . Hence, pump efficiency hpump ! 1, implying that water horsepower Wwater horsepower and brake horsepower bhp are the same. Using Eqs. 14–3 and 14–4, # # # bhp ! vTshaft ! rvV (r2V2, t # r1V1, t) ! W water horsepower ! rgV H
(14–16)
which is solved for net head H, 1 H ! (vr2V2, t # vr1V1, t) g
EXAMPLE 14–5
V2, n
(14–14)
Or, in terms of angles a1 and a2 and the magnitudes of the absolute velocity vectors,
Net head:
→
# Tshaft ! rV (r2V2, t # r1V1, t)
(14–17)
Idealized Blower Performance
. A centrifugal blower rotates at n ! 1750 rpm (183.3 rad/s). Air enters the impeller normal to the blades (a1 ! 0°) and exits at an angle of 40° from radial (a2 ! 40°) as sketched in Fig. 14–38. The inlet radius is r1 ! 4.0 cm, and the inlet blade width b1 ! 5.2 cm. The outlet radius is r2 ! 8.0 cm, and the outlet blade width b2 ! 2.3 cm. The volume flow rate is 0.13 m3/s. Assuming 100 percent efficiency, calculate the net head produced by this blower in equivalent millimeters of water column height. Also calculate the required brake horsepower in watts.
SOLUTION We are to calculate the brake horsepower and net head of an idealized blower at a given volume flow rate and rotation rate. Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between rotor blades and blower casing. 3 The air is incompressible. 4 The efficiency of the blower is 100 percent (no irreversible losses). Properties We take the density of air to be rair ! 1.20 kg/m3. Analysis Since the volume flow rate (capacity) is given, we calculate the normal velocity components at the inlet and the outlet using Eq. 14–12, # V 0.13 m3/s V1, n ! ! 9.947 m/s ! 2pr1b1 2p(0.040 m)(0.052 m)
(1)
V1 ! V1, n, and V1, t ! 0, since a1 ! 0°. Similarly, V2, n ! 11.24 m/s, and
V2, t ! V2, n tan a 2 ! (11.24 m/s) tan(40() ! 9.435 m/s
(2)
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759 CHAPTER 14
Now we use Eq. 14–17 to predict the net head,
H!
(3)
F
v 183.3 rad/s (r V # r1V1, t ) ! (0.080 m)(9.435 m/s) ! 14.1 m g 2 2, t 9.81 m/s2 0
Note that the net head of Eq. 3 is in meters of air, the pumped fluid. To convert to pressure in units of equivalent millimeters of water column, we multiply by the ratio of air density to water density,
Hwater column ! H
r air r water
! (14.1 m)
1.20 kg/m3 1000 mm a b ! 17.0 mm of water 1m 998 kg/m3
(4)
Finally, we use Eq. 14–16 to predict the required brake horsepower,
# W's bhp ! rgV H ! (1.20 kg/m3)(9.81 m/s2)(0.13 m3/s)(14.1 m)a b kg ' m/s2 ! 21.6 W
(5)
Discussion Note the unit conversion in Eq. 5 from kilograms, meters, and seconds to watts; this conversion turns out to be useful in many turbomachinery calculations. The actual net head delivered to the air will be lower than that predicted by Eq. 3 due to inefficiencies. Similarly, actual brake horsepower will be higher than that predicted by Eq. 5 due to inefficiencies in the blower, friction on the shaft, etc.
In order to design the shape of the impeller blades, we must use trigonometry to obtain expressions for V1, t and V2, t in terms of blade angles b1 and 14–39) to the triangle in→ Fig. 14–36 b2. Applying the law of cosines (Fig. → formed by absolute velocity vector V 2, relative velocity vector V 2, relative, and the tangential velocity of the blade at radius r2 (of magnitude vr2) we get V 22
!
V 22, relative
$
v 2r 22
# 2vr2V2, relative cos b 2
(14–18)
Law of Cosines b A
C a B
c
c2 = a2 + b2 – 2ab cos C
But we also see from Fig. 14–36 that V2, relative cos b 2 ! vr2 # V2, t
FIGURE 14–39 The law of cosines is utilized in the analysis of a centrifugal pump.
Substitution of this equation into Eq. 14–18 yields 1 vr2V2, t ! (V 22 # V 22, relative $ v 2r 22) 2
(14–19)
A similar equation results for the blade inlet (change all subscripts 2 in Eq. 14–19 to subscript 1). Substitution of these into Eq. 14–17 yields Net head:
H!
1 [(V 22 # V 21) $ (v 2r 22 # v 2r 21) # (V 22, relative # V 21, relative)] 2g
(14–20)
In words, Eq. 14–20 states that in the ideal case (no irreversible losses), the net head is proportional to the change in absolute kinetic energy plus the rotor-tip kinetic energy change minus the change in relative kinetic energy from inlet to outlet of the impeller. Finally, equating Eq. 14–20 and Eq. 14–2,
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760 FLUID MECHANICS
where we set subscript 2 as the outflow and subscript 1 as the inflow, we see that
→
Vrelative
→
V
a
r
r
v Absolute
v Rotating
FIGURE 14–40 For the approximation of flow through an impeller with no irreversible losses, it is often more convenient to work with a relative frame of reference rotating with the impeller; in that case, the Bernoulli equation gets an additional term, as indicated in Eq. 14–22. →
V1, relative V1, t b1
V1, n
→
V1
V 2relative v 2r 2 V 2relative v 2r 2 P P $ $ # $ zb ! a # $ zb rg rg 2g 2g 2g 2g out in
Note that we are not limited to analysis of only the inlet and outlet. In fact, we may apply Eq. 14–21 to any two radii along the impeller. In general then, we write an equation that is commonly called the Bernoulli equation in a rotating reference frame: V 2relative v 2r 2 P $ # $ z ! constant rg 2g 2g
vr1
V1, t ! vr1 #
V1, n tan b 1
(14–23)
A similar expression is obtained for V2, t (replace subscript 1 by 2), or in fact for any radius between r1 and r2. When V1, t ! 0 and V1, n ! V1,
v
FIGURE 14–41 Close-up frontal view of the velocity vectors at the impeller blade inlet. The absolute velocity vector is shown as a bold arrow.
(14–22)
We see that Eq. 14–22 is the same as the usual Bernoulli equation, except that since the speed used is the relative speed (in the rotating reference frame), an “extra” term (the third term on the left of Eq. 14–22) appears in the equation to account for rotational effects (Fig. 14–40). We emphasize that Eq. 14–22 is an approximation, valid only for the ideal case in which there are no irreversible losses through the impeller. Nevertheless, it is valuable as a first-order approximation for flow through the impeller of a centrifugal pump. We now examine Eq. 14–17, the equation for net head, more closely. Since the term containing V1, t carries a negative sign, we obtain the maximum H by setting V1, t to zero. (We are assuming that there is no mechanism in the eye of the pump that can generate a negative value of V1, t.) Thus, a first-order approximation for the design condition of the pump is to set V1, t ! 0. In other words, we select the blade inlet angle b1 such that the flow into the impeller blade is purely radial from an absolute reference frame, and V1, n ! V1. The velocity vectors at r ! r1 in Fig. 14–36 are magnified and redrawn in Fig. 14–41. Using some trigonometry we see that
b1 r1
(14–21)
vr1 !
V1, n tan b 1
(14–24)
Finally, combining Eq. 14–24 with Eq. 14–12, we have an expression for volume flow rate as a function of inlet blade angle b1 and rotational speed, # V ! 2pb1vr 21 tan b 1
(14–25)
Equation 14–25 can be used for preliminary design of the impeller blade shape as illustrated by Example 14–6. EXAMPLE 14–6
Preliminary Design of a Centrifugal Pump
A centrifugal pump is being designed to pump liquid refrigerant R-134a at room temperature and atmospheric pressure. The impeller inlet and outlet radii are r1 ! 100 and r2 ! 180 mm, respectively (Fig. 14–42). The impeller inlet and outlet widths are b1 ! 50 and b2 ! 30 mm (into the page
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761 CHAPTER 14
of Fig. 14–42). The pump is to deliver 0.25 m3/s of the liquid at a net head of 14.5 m when the impeller rotates at 1720 rpm. Design the blade shape for the case in which these operating conditions are the design conditions of the pump (V1, t ! 0, as sketched in the figure); specifically, calculate angles b1 and b2, and discuss the shape of the blade. Also predict the horsepower required by the pump.
SOLUTION For a given flow rate, net head, and dimensions of a centrifugal pump, we are to design the blade shape (leading and trailing edge angles). We are also to estimate the horsepower required by the pump. Assumptions 1 The flow is steady. 2 The liquid is incompressible. 3 There are no irreversible losses through the impeller. 4 This is only a preliminary design. Properties For refrigerant R-134a at T ! 20°C, vf ! 0.0008157 m3/kg. Thus r ! 1/vf ! 1226 kg/m3. Analysis We calculate the required water horsepower from Eq. 14–3,
# # Wwater horsepower ! rgV H
V2, n
→
V2
vr2 →
b2
V1, relative
→
b2
V2, relative
→
V1
b1 r2
b1 v
r1
vr1
FIGURE 14–42 Relative and absolute velocity vectors and geometry for the centrifugal pump impeller design of Example 14–6.
W's b ! (1226 kg/m3)(9.81 m/s2)(0.25 m3/s)(14.5 m)a kg ' m/s2 ! 43,600 W The required brake horsepower will be greater than this in a real pump. However, in keeping with the approximations for this preliminary design, we assume 100 percent efficiency such that bhp is approximately equal to . Wwater horsepower,
# hp b ! 58.5 hp bhp ! W water horsepower ! 43,600 Wa 745.7 W
We report the final result to two significant digits in keeping with the precision of the given quantities; thus, bhp " 59 horsepower. In all calculations with rotation, we need to convert the rotational speed . from n (rpm) to v (rad/s), as illustrated in Fig. 14–43,
v ! 1720
rot 2p rad 1 min a ba b ! 180.1 rad/s min rot 60 s
CAUTION Always convert rotation rate from rpm to radians per second.
(1)
We calculate the blade inlet angle using Eq. 14–25,
# V 0.25 m3/s b ! arctan a b ! 23.8( b 1 ! arctan a 2 2pb1vr 1 2p(0.050 m)(180.1 rad/s)(0.10 m)2
We find b2 by utilizing the equations derived earlier for our elementary analysis. First, for the design condition in which V1, t ! 0, Eq. 14–17 reduces to
Net head:
F
vr2V2, t 1 H ! (vr2V2, t # vr1V1, t ) ! g g 0
from which we calculate the tangential velocity component,
V2, t !
gH vr2
(2)
Using Eq. 14–12, we calculate the normal velocity component,
V2, n !
# V 2pr2b2
(3)
FIGURE 14–43 Proper unit conversion requires the units of rotation rate to be rad/s.
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762 FLUID MECHANICS
Next, we perform the same trigonometry used to derive Eq. 14–23, but on the trailing edge of the blade rather than the leading edge. The result is
V2, t ! vr2 #
V2, n tan b 2
from which we finally solve for b2,
V2, n b 2 ! arctan a b vr2 # V2,t
(4)
After substitution of Eqs. 2 and 3 into Eq. 4, and insertion of the numerical values, we obtain
b 2 ! 14.7(
b2 b2 b1 b2 r2
v
r1
FIGURE 14–44 Three possible blade shapes for the centrifugal pump impeller design of Example 14–6. All three blades have leading edge angle b1 ! 24° and trailing edge angle b2 ! 15°, but differ in how b is varied with the radius. The drawing is to scale.
We report the final results to only two significant digits. Thus our preliminary design requires backward-inclined impeller blades with b1 ≅ 24° and b2 ≅ 15°. Once we know the leading and trailing edge blade angles, we design the detailed shape of the impeller blade by smoothly varying blade angle b from b1 to b2 as radius increases from r1 to r2. As sketched in Fig. 14–44, the blade can be of various shapes while still keeping b1 ≅ 24° and b2 ≅ 15°, depending on how we vary b with the radius. In the figure, all three blades begin at the same location (zero absolute angle) at radius r1; the leading edge angle for all three blades is b1 ! 24°. The medium length blade (the gray one in Fig. 14–44) is constructed by varying b linearly with r. Its trailing edge intercepts radius r2 at an absolute angle of approximately 93°. The longer blade (the black one in the figure) is constructed by varying b more rapidly near r1 than near r2. In other words, the blade curvature is more pronounced near its leading edge than near its trailing edge. It intercepts the outer radius at an absolute angle of about 114°. Finally, the shortest blade (the blue blade in Fig. 14–44) has less blade curvature near its leading edge, but more pronounced curvature near its trailing edge. It intercepts r2 at an absolute angle of approximately 77°. It is not immediately obvious which blade shape is best. Discussion Keep in mind that this is a preliminary design in which irreversible losses are ignored. A real pump would have losses, and the required brake horsepower would be higher (perhaps 20 to 30 percent higher) than the value estimated here. In a real pump with losses, a shorter blade has less skin friction drag, but the normal stresses on the blade are larger because the flow is turned more sharply near the trailing edge where the velocities are largest; this may lead to structural problems if the blades are not very thick, especially when pumping dense liquids. A longer blade has higher skin friction drag, but lower normal stresses. In addition, you can see from a simple blade volume estimate in Fig. 14–44 that for the same number of blades, the longer the blades, the more flow blockage, since the blades are of finite thickness. In addition, the displacement thickness effect of boundary layers growing along the blade surfaces (Chap. 10) leads to even more pronounced blockage for the long blades. Obviously some engineering optimization is required to determine the exact shape of the blade.
How many blades should we use in an impeller? If we use too few blades, circulatory flow loss will be high. Circulatory flow loss occurs because there is a finite number of blades. Recall that in our preliminary analysis,
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763 CHAPTER 14
we assume a uniform tangential velocity V2, t around the entire circumference of the outlet of the control volume (Fig. 14-37). This is strictly correct only if we have an infinite number of infinitesimally thin blades. In a real pump, of course, the number of blades is finite, and the blades are not infinitesimally thin. As a result, the tangential component of the absolute velocity vector is not uniform, but drops off in the spaces between blades as illustrated in Fig. 14-45a. The net result is an effectively smaller value of V2, t, which in turn decreases the actual net head. This loss of net head (and pump efficiency) is called circulatory flow loss. On the other hand, if we have too many blades (as in Fig. 14-45b) there will be excessive flow blockage losses and losses due to the growing boundary layers, again leading to nonuniform flow speeds at the outer radius of the pump and lower net head and efficiency. These losses are called passage losses. The bottom line is that some engineering optimization is necessary in order to choose both the blade shape and number of blades. Such analysis is beyond the scope of the present text. A quick perusal through the turbomachinery literature shows that 11, 14, and 16 are common numbers of rotor blades for medium-sized centrifugal pumps. Once we have designed the pump for specified net head and flow rate (design conditions), we can estimate its net head at conditions away from design conditions. In other words, keeping b1, b2, r1, r2, b1, b2, and v fixed, we vary the volume flow rate above and below the design flow rate. We have all the equations: Eq. 14–17 for net head H in terms of absolute tangential velocity components V1, t and V2, t, Eq. 14–23 for V1, t and V2, t as functions of absolute normal velocity components V1, n and . V2, n, and Eq. 14–12 for V1, n and V2, n as functions of volume flow rate V. In Fig. 14–46
V2, t
v
(a)
v
(b)
FIGURE 14-45 (a) A centrifugal pump impeller with too few blades leads to excessive circulatory flow loss—the tangential velocity at outer radius r2 is smaller in the gaps between blades than at the trailing edges of the blades (absolute tangential velocity vectors are shown). (b) On the other hand, since real impeller blades have finite thickness, an impeller with too many blades leads to passage losses due to excessive flow blockage and large skin friction drag (velocity vectors in a frame of reference rotating with the impeller are shown exiting one blade row). The bottom line is that pump engineers must optimize both blade shape and number of blades.
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764 FLUID MECHANICS 30
Irreversible losses
25
Predicted performance
20
Actual performance
H, m 15 10 5
Design conditions
0 0.2
0.25 • V, m3/s
0.3
FIGURE 14-46 Net head as a function of volume flow rate for the pump of Example 14-6. The difference between predicted and actual performance is due to unaccounted irreversibilities in the prediction.
Axial Pumps
FL
Streamlines
FIGURE 14-47 The blades of an axial-flow pump behave like the wing of an airplane. The air is turned downward by the wing as it generates lift force FL.
Downwash
. we combine these equations to generate a plot of H versus V for the pump designed in Example 14–6. The solid blue line is the predicted performance, based on our preliminary analysis. The predicted performance curve is . nearly linear with V both above and below design conditions since the term vr1V1, t in Eq. 14–17 is small compared to the term vr2V2, t. Recall that at the predicted design conditions, we had set V1, t ! 0. For volume flow rates higher than this, V1, t is predicted by Eq. 14–23 to be negative. In keeping with our previous assumptions, however, it is not possible to have negative values of V1, t. Thus, the slope of the predicted performance curve changes suddenly beyond the design conditions. Also sketched in Fig. 14–46 is the actual performance of this centrifugal pump. While the predicted performance is close to the actual performance at design conditions, the two curves deviate substantially away from design conditions. At all volume flow rates, the actual net head is lower than the predicted net head. This is due to irreversible effects such as friction along blade surfaces, leakage of fluid between the blades and the casing, prerotation (swirl) of fluid in the region of the eye, flow separation on the leading edges of the blades (shock losses) or in the expanding portions of the flow passages, circulatory flow loss, passage loss, and irreversible dissipation of swirling eddies in the volute, among other things.
Lift
Low P High P
FIGURE 14-48 Downwash and pressure rise across the rotor plane of a helicopter, which is a type of axial-flow pump.
Axial pumps do not utilize so-called centrifugal forces. Instead, the impeller blades behave more like the wing of an airplane (Fig. 14–47), producing lift by changing the momentum of the fluid as they rotate. The rotor of a helicopter, for example, is a type of axial-flow pump (Fig. 14–48). The lift force on the blade is caused by pressure differences between the top and bottom surfaces of the blade, and the change in flow direction leads to downwash (a column of descending air) through the rotor plane. From a time-averaged perspective, there is a pressure jump across the rotor plane that induces a downward airflow (Fig. 14–48). Imagine turning the rotor plane vertically; we now have a propeller (Fig. 14–49a). Both the helicopter rotor and the airplane propeller are examples of open axial-flow fans, since there is no duct or casing around the tips of the blades. The common window fan you install in your bedroom window in the summer operates under the same principles, but the goal is to blow air rather than to provide a force. Be assured, however, that there is a net force acting on the fan housing. If air is blown from left to right, the force on the fan acts to the left, and the fan is held down by the window sash. The casing around the house fan also acts as a short duct, which helps to direct the flow and eliminate some losses at the blade tips. The small cooling fan inside your computer is typically an axial-flow fan; it looks like a miniature window fan (Fig. 14–49b) and is an example of a ducted axial-flow fan. If you look closely at the airplane propeller blade in Fig. 14–49a, the rotor blade of a helicopter, the propeller blade of a radio-controlled model airplane, or even the blade of a well-designed window fan, you will notice some twist in the blade. Specifically, the airfoil at a cross section near the hub or root of the blade is at a higher pitch angle (u) than the airfoil at a
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765 CHAPTER 14
FIGURE 14–49 Axial-flow fans may be open or ducted: (a) a propeller is an open fan, and (b) a computer cooling fan is a ducted fan. (a)
(a) Courtesy Whirl Wind Propellers Corporation. Used by permission. (b) Courtesy ebm-papst Mulfingen GmbH & Co. KG. Used by permission.
(b)
cross section near the tip, uroot " utip (Fig. 14–50). This is because the tangential speed of the blade increases linearly with radius, u u ! vr
(14–26)
Wind direction
v
→
Vwind
→
At a given radius then, the velocity V relative of the air relative to→ the blade is estimated to first order as→ the vector sum of wind velocity V wind and the negative of blade velocityV blade, →
→
→
Vrelative ! Vwind # Vblade
(14–27)
uroot Hub Root
uu = vr
→
where the magnitude of V blade is equal →to the tangential blade speed uu, as given by Eq. 14-26. The direction of V blade is tangential to the →rotational path of the blade. At the blade position sketched in Fig. 14–50, V blade is to the left. → In Fig. 14–51 we compute V relative graphically using Eq. 14–27 at two radii—the root radius and the tip radius of the rotor blade sketched in Fig. 14–50. As you can see, the relative angle of attack a is the same in either case. In fact, the amount of twist is determined by setting pitch angle u such that a is the same at any radius. → Note also that the magnitude of the relative velocity V relative increases from the root to the tip. It follows that the dynamic pressure encountered by cross sections of the blade increases with radius, and the lift force per unit width into the page in Fig. 14–51 also increases with radius. Propellers tend to be narrower at the root and wider toward the tip in order to take advantage of the larger lift contribution available toward the tip. At the very tip, however, the blade is usually rounded off to avoid excessive induced drag (Chap. 11) that would exist if the blade were simply chopped off abruptly as in Fig. 14–50. Equation 14–27 is not exact for several reasons. First, the rotating motion of the rotor introduces some swirl to the airflow (Fig. 14–52). This reduces
utip Tip
FIGURE 14–50 A well-designed rotor blade or propeller blade has twist, as shown by the blue cross-sectional slices through one of the three blades; blade pitch angle u is higher at the root than at the tip because the tangential speed of the blade increases with radius.
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766 FLUID MECHANICS
a →
a
Vrelative
utip
→
Vblade
→
Vwind
→
Vwind
FIGURE 14–51 Graphical computation of vector → V relative at two radii: (a) root, and (b) tip of the rotor blade sketched in Fig. 14–50.
FIGURE 14–52 The rotating blades of a rotor or propeller induce swirl in the surrounding fluid.
Dpropeller
→
→
Vblade
Vrelative
(a)
(b)
the effective tangential speed of the blade relative to the wind. Second, since the hub of the rotor is of finite size, the air accelerates around it, causing the wind speed to increase locally at cross sections of the blade close to the root. Third, the axis of the rotor or propeller may not be aligned exactly parallel to the wind. Finally, the wind speed itself is not easily determined because it turns out that the wind accelerates as it approaches the whirling rotor. There are methods available to approximate these and other secondary effects, but they are beyond the scope of the present text. The first-order approximation given by Eq. 14–27 is adequate for preliminary rotor and propeller design, as illustrated in Example 14–7.
v
Vwind
→
uroot
Airplane nose
D hub v
→
Vblade
FIGURE 14–53 Setup for the design of the model airplane propeller of Example 14–7, not to scale.
EXAMPLE 14–7
Calculation of Twist in an Airplane Propeller
Suppose you are designing the propeller of a radio-controlled model airplane. The overall diameter of the propeller is 34.0 cm, and the hub assembly diameter is 5.5 cm (Fig. 14–53). The propeller rotates at 1700 rpm, and the airfoil chosen for the propeller cross section achieves its maximum efficiency at an angle of attack of 14°. When the airplane flies at 30 mi/h (13.4 m/s), calculate the blade pitch angle from the root to the tip of the blade such that a ! 14° everywhere along the propeller blade.
SOLUTION We are to calculate blade pitch angle u from the root to the tip of the propeller such that the angle of attack is a ! 14° at every radius along the propeller blade. Assumptions 1 The air at these low speeds is incompressible. 2 We neglect the secondary effects of swirl and →acceleration of the air as it approaches the propeller; i.e., the magnitude of Vwind is assumed to equal the speed of the aircraft. 3 The airplane flies level, such that the propeller axis is parallel to the wind velocity. Analysis The velocity of the air relative to the blade is approximated to first order at any radius by using Eq. 14–27. A sketch of the velocity vectors at some arbitrary radius r is shown in Fig. 14–54. From the geometry we see that Pitch angle at arbitrary radius r:
u!a$f
(1)
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767 CHAPTER 14
and
→
0 Vwind 0
→
0 Vwind 0 ! arctan f ! arctan → vr 0 Vblade 0
(2)
where we have also used Eq. 14–26 for the blade speed at radius r. At the root (r ! Dhub/2 ! 2.75 cm), Eq. 2 becomes
13.4 m/s 1 rot 60 s a ba b d ! 83.9( u ! a $ f ! 14( $ arctan c (1700 rot/min)(0.0275 m) 2p rad min
Similarly, the pitch angle at the tip (r ! Dpropeller/2 ! 17.0 cm) is
13.4 m/s 1 rot 60 s u ! a $ f ! 14( $ arctan c a ba b d ! 37.9( (1700 rot/min)(0.17 m) 2p rad min
u
At radii between the root and the tip, Eqs. 1 and 2 are used to calculate u as a function of r. Results are plotted in Fig. 14–55. Discussion The pitch angle is not linear because of the arctangent function in Eq. 2.
→
Vrelative
→
–Vblade
→
Vwind
FIGURE 14–54 Velocity vectors at some arbitrary radius r of the propeller of Example 14–7. 90 80 70 u, degrees
Airplane propellers have variable pitch, meaning that the pitch of the entire blade can be adjusted by rotating the blades through mechanical linkages in the hub. For example, when a propeller-driven airplane is sitting at the airport, warming up its engines at high rpm, why does it not start moving? Well, for one thing, the brakes are being applied. But more importantly, propeller pitch is adjusted so that the average angle of attack of the airfoil cross sections is zero—no net thrust is provided. While the airplane taxies to the runway, the pitch is adjusted so as to produce a small amount of thrust. As the plane takes off, the engine rpm is high, and the blade pitch is adjusted such that the propeller delivers the maximum thrust. In most cases the pitch can even be adjusted “backward” (negative angle of attack) to provide reverse thrust to slow down the airplane after landing. We plot qualitative performance curves for a typical propeller fan in Fig. 14–56. Unlike centrifugal fans, brake horsepower tends to decrease with flow rate. In addition, the efficiency curve leans more to the right compared to that of centrifugal fans (see Fig. 14–8). The result is that efficiency drops off rapidly for volume flow rates higher than that at the best efficiency point. The net head curve also decreases continuously with flow rate (although there are some wiggles), and its shape is much different than that of a centrifugal flow fan. If the head requirements are not severe, propeller fans can be operated beyond the point of maximum efficiency to. achieve higher volume flow rates. Since bhp decreases at high values of V, there is not a power penalty when the fan is run at high flow rates. For this reason it is tempting to install a slightly undersized fan and push it beyond its best efficiency point. At the other extreme, if operated below its maximum efficiency point, the flow may be noisy and unstable, which indicates that the fan may be oversized (larger than necessary). For these reasons, it is usually best to run a propeller fan at, or slightly above, its maximum efficiency point.
f
a
Hub
60 Tip 50 40 30 0
5
10 r, cm
15
20
FIGURE 14–55 Blade pitch angle as a function of radius for the propeller of Example 14–7.
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768 FLUID MECHANICS
hpump
H, hpump, or bhp
H
bhp
0
⋅ V
0
FIGURE 14–56 Typical fan performance curves for a propeller (axial-flow) fan.
Impeller
Hub
Motor
v
(a) Impeller 1
Hub
Motor
v v Impeller 2 Impeller
Gear box (b) Hub
Motor
v v Stator (c)
FIGURE 14–57 A tube-axial fan (a) imparts swirl to the exiting fluid, while (b) a counterrotating axial-flow fan and (c) a vaneaxial fan are designed to remove the swirl.
When used to move flow in a duct, a single-impeller axial-flow fan is called a tube-axial fan (Fig. 14–57a). In many practical engineering applications of axial-flow fans, such as exhaust fans in kitchens, building ventilation duct fans, fume hood fans, and automotive radiator cooling fans, the swirling flow produced by the rotating blades (Fig. 14–57a) is of no concern. But the swirling motion and increased turbulence intensity can continue for quite some distance downstream, and there are applications where swirl (or its affiliated noise and turbulence) is highly undesirable. Examples include wind tunnel fans, torpedo fans, and some specialized mine shaft ventilation fans. There are two basic designs that largely eliminate swirl: A second rotor that rotates in the opposite direction can be added in series with the existing rotor to form a pair of counter-rotating rotor blades; such a fan is called a counter-rotating axial-flow fan (Fig. 14–57b). The swirl caused by the upstream rotor is cancelled by an opposite swirl caused by the downstream rotor. Alternatively, a set of stator blades can be added either upstream or downstream of the rotating impeller. As implied by their name, stator blades are stationary (nonrotating) guide vanes that simply redirect the fluid. An axial-flow fan with a set of rotor blades (the impeller or the rotor) and a set of stator blades called vanes (the stator) is called a vane-axial fan (Fig. 14–57c). The stator blade design of the vane-axial fan is much simpler and less expensive to implement than is the counter-rotating axial-flow fan design. The swirling fluid downstream of a tube-axial fan wastes kinetic energy and has a high level of turbulence; the vane-axial fan partially recovers this wasted kinetic energy and reduces the level of turbulence. Vane-axial fans are thus both quieter and more energy efficient than tube-axial fans. A properly designed counter-rotating axial-flow fan may be even quieter and more energy efficient. Furthermore, since there are two sets of rotating blades, a higher pressure rise can be obtained with the counter-rotating design. The construction of a counter-rotating axial-flow fan is more complex, of course, requiring either two synchronized motors or a gear box. Axial-flow fans can be either belt driven or direct drive. The motor of a direct-drive vane-axial fan is mounted in the middle of the duct. It is common practice (and good design) to use the stator blades to provide physical support for the motor. Photographs of a belt-driven tube-axial fan and a direct-drive vane-axial fan are provided in Fig. 14–58. The stator blades of the vane-axial fan can be seen behind (downstream of) the rotor blades in Fig. 14–58b. An alternative design is to place the stator blades upstream of the impeller, imparting preswirl to the fluid. The swirl caused by the rotating impeller blades then removes this preswirl. It is fairly straightforward to design the shape of the blades in all these axial-flow fan designs, at least to first order. For simplicity, we assume thin blades (e.g., blades made out of sheet metal) rather than airfoil-shaped blades. Consider, for example, a vane-axial flow fan with rotor blades upstream of stator blades (Fig. 14–59). The distance between the rotor and stator has been exaggerated in this figure to enable velocity vectors to be drawn between the blades. The hub radius of the stator is assumed to be the same as the hub radius of the rotor so that the cross-sectional area of flow remains constant. As we did previously with the propeller, we consider the cross section of one impeller blade as it passes vertically in front of us.
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769 CHAPTER 14
FIGURE 14–58 Axial-flow fans: (a) a belt-driven tube-axial fan without stator blades, and (b) a direct-drive vane-axial fan with stator blades to reduce swirl and improve efficiency. (a)
(b)
Since there are multiple blades, the next blade passes by shortly thereafter. At a chosen radius r, we make the two-dimensional approximation that the blades pass by as an infinite series of two-dimensional blades called a blade row or cascade. A similar assumption is made for the stator blades, even though they are stationary. Both blade rows are sketched in Fig. 14–59. In Fig. 14–59b, the velocity vectors are seen from an absolute reference frame, i.e., that of a fixed observer looking horizontally at the vane-axial flow fan. Flow enters from the left at speed Vin in the horizontal (axial) direction. The rotor blade row moves at constant speed vr vertically upward in this reference frame, as indicated. Flow is turned by these moving blades and leaves the trailing edge upward and to the right as indicated in Fig. → indicates rotor trailing edge.) 14–59b as vector V rt. (The subscript notation → To find the magnitude and direction of V rt, we redraw the blade rows and vectors in a relative reference frame (the frame of reference of the rotating rotor blade) in Fig. 14–59c. This reference frame is obtained by subtracting the rotor blade velocity (adding a vector of magnitude vr pointing vertically downward) from all velocity vectors. As shown in Fig. 14–59c, the velocity → vector relative to the leading edge of the rotor blade is V in, relative, calculated → as the vector sum of V in and the downward→ vector of magnitude vr. We adjust the pitch of the rotor blade such that V in, relative is parallel (tangential) to the leading edge of the rotor blade at this cross section. Flow is turned by the rotor blade. We assume that the flow leaving the rotor blade is parallel to the blade’s trailing edge→ (from the relative reference also know that frame), as sketched in Fig. 14–59c →as vector V rt, relative. We → the horizontal (axial) component of V rt, relative must equal V in in order to conserve mass. Note that we are assuming incompressible flow and constant flow area normal to the page in Fig. 14–59. Thus, the axial component of velocity must be everywhere equal to Vin. This piece of information estab→ lishes the→ magnitude of vector V rt, relative, which is not the same as the magniReturning to the absolute reference frame of Fig. 14–59b, tude of V in, relative. → → absolute velocity V rt is calculated as the vector sum of V rt, relative and the vertically upward vector of magnitude vr.
(a) © Barry Blower, ASC LP. Used by permission. (b) Photo courtesy of Howden Buffalo, Inc. Used by permission.
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770 FLUID MECHANICS →
v vr →
→
Vin
Vout r
Rotor blade row
Stator blade row (a)
vr →
→
Vrt
Vout
→
Vin
→
Vrt,relative (b)
→
Vin →
Vin,relative
vr
→
vr
Vout
vr →
Vin →
Vrt,relative
→
Vout,relative (c)
FIGURE 14–59 Analysis of a vane-axial flow fan at radius r using the two-dimensional blade row approximation; (a) overall view, (b) absolute reference frame, and (c) reference frame relative to the rotating rotor blades (impeller).
Finally, the stator blade is designed such that V rt is parallel to the leading edge of the stator blade. The flow is once again turned, this time by the stator blade. Its trailing edge is horizontal so that the flow leaves axially (without any swirl). The final outflow velocity must be identical to the inflow velocity by conservation of mass if we assume incompressible flow and → → constant flow area normal to the page. In other words, V out ! V in. For completeness, the outflow velocity→in the relative reference frame is sketched in → Fig. 14–59c. We also see that V out, relative ! V in, relative. Now imagine repeating this analysis for all radii from the hub to the tip. As with the propeller, we would design our blades with some twist since the value of vr increases with radius. A modest improvement in efficiency can be gained at design conditions by using airfoils instead of sheet metal blades; the improvement is more significant at off-design conditions. If there are, say, seven rotor blades in a vane-axial fan, how many stator blades should there be? You might at first say seven so that the stator matches the rotor—but this would be a very poor design! Why? Because at the instant in time when one blade of the rotor passes directly in front of a stator blade, all six of its brothers would do the same. Each stator blade would simultaneously encounter the disturbed flow in the wake of a rotor blade. The resulting flow would be both pulsating and noisy, and the entire unit would vibrate severely. Instead, it is good design practice to choose the number of stator blades such that it has no common denominator with the number of rotor blades. Combinations like seven and eight, seven and nine, six and seven, or nine and eleven are good choices. Combinations like eight and ten (common denominator of two) or nine and twelve (common denominator of three) are not good choices. We plot the performance curves of a typical vane-axial flow fan in Fig. 14–60. The general shapes are very similar to those of a propeller fan (Fig. 14–56), and you are referred to the discussion there. After all, a vane-axial flow fan is really the same as a propeller fan or tube-axial flow fan except for the additional stator blades that straighten the flow and tend to smooth out the performance curves. As discussed previously, an axial-flow fan delivers high volume flow rate, but fairly low pressure rise. Some applications require both high flow rate and high pressure rise. In such cases, several stator–rotor pairs can be combined in series, typically with a common shaft and common hub (Fig. 14–61). When two or more rotor–stator pairs are combined like this we call it a multistage axial-flow pump. A blade row analysis similar to the one of Fig. 14–59 is applied to each successive stage. The details of the analysis can get complicated, however, because of compressibility effects and because the flow area from the hub to the tip may not remain constant. In a multistage axial-flow compressor, for example, the flow area decreases downstream. The blades of each successive stage get smaller as the air gets further compressed. In a multistage axial-flow turbine, the flow area typically grows downstream as pressure is lost in each successive stage of the turbine. One well-known example of a turbomachine that utilizes both multistage axial-flow compressors and multistage axial-flow turbines is the turbofan engine used to power modern commercial airplanes. A cutaway schematic diagram of a turbofan engine is shown in Fig. 14–62. Some of the air passes through the fan, which delivers thrust much like a propeller. The rest of the
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771 CHAPTER 14
EXAMPLE 14–8
Design of a Vane-Axial Flow Fan for a Wind Tunnel
A vane-axial flow fan is being designed to power a wind tunnel. There must not be any swirl in the flow downstream of the fan. It is decided that the stator blades should be upstream of the rotor blades (Fig. 14–63) to protect the impeller blades from damage by objects that might accidentally get blown into the fan. To reduce expenses, both the stator and rotor blades are to be constructed of sheet metal. The leading edge of each stator blade is aligned axially (bsl ! 0.0°) and its trailing edge is at angle bst ! 60.0° from the axis as shown in the sketch. (The subscript notation “sl” indicates stator leading edge and “st” indicates stator trailing edge.) There are 16 stator blades. At design conditions, the axial-flow speed through the blades is 47.1 m/s, and the impeller rotates at 1750 rpm. At radius r ! 0.40 m, calculate the leading and trailing edge angles of the rotor blade, and sketch the shape of the blade. How many rotor blades should there be?
SOLUTION For given flow conditions and stator blade shape at a given radius, we are to design the rotor blade. Specifically, we are to calculate the leading and trailing edge angles of the rotor blade and sketch its shape. We are also to decide how many rotor blades to construct.
hpump
H H, hpump, or bhp
air passes through a low-pressure compressor, a high-pressure compressor, a combustion chamber, a high-pressure turbine, and then finally a low-pressure turbine. The air and products of combustion are then exhausted at high speed to provide even more thrust. Computational fluid dynamics (CFD) codes are obviously quite useful in the design of such complex turbomachines (Chap. 15).
bhp
0
⋅ V
0
FIGURE 14–60 Typical fan performance curves for a vane-axial flow fan. Rotor 1
Rotor 2
v
Stator 1
Rotating hub
Shaft
Stator 2
FIGURE 14–61 A multistage axial-flow pump consists of two or more rotor–stator pairs.
FIGURE 14–62 Pratt & Whitney PW4000 turbofan engine; an example of a multistage axial-flow turbomachine. Courtesy Pratt & Whitney. Used by permission.
cen72367_ch14.qxd 11/12/04 2:11 PM Page 772
772 FLUID MECHANICS
Hub and motor vr
bst ?? r
→
Vin Stator
v
→
Vout
Assumptions 1 The air is nearly incompressible. 2 The flow area between the hub and tip is constant. 3 Two-dimensional blade row analysis is appropriate. Analysis First we analyze flow through the stator from an absolute reference frame, using the two-dimensional approximation of a cascade (blade row) of stator blades (Fig. 14–64). Flow enters axially (horizontally) and is turned 60.0° downward. Since the axial component of velocity must remain constant to conserve mass, the magnitude of the velocity leaving the trailing → edge of the stator, V st, is calculated as
Vst !
Rotor
FIGURE 14–63 Schematic diagram of the vane-axial flow fan of Example 14–8. The stator precedes the rotor, and the shape of the rotor blade is unknown—it is to be designed.
Vin 47.1 m/s ! 94.2 m/s ! cos b st cos (60.0()
(1)
→
The direction of V st is assumed to be that of the stator trailing edge. In other words, we assume that the flow turns nicely through the blade row and exits parallel to the trailing edge of the blade, as shown in Fig. 14–64. → We convert V st to the relative reference frame moving with the rotor blades. At a radius of 0.40 m, the tangential velocity of the rotor blades is
2p rad 1 min ba b (0.40 m) ! 73.30 m/s rot 60 s
u u ! vr ! (1750 rot/min)a
(2)
Since the rotor blade row moves upward in Fig. 14–63, we add a downward → velocity with magnitude given by Eq. 2 to translate V st into the rotating reference frame sketched in Fig. 14–65. The angle of the leading edge of the rotor, brl, is calculated by using trigonometry,
b rl ! arctan ! arctan →
Vin
bst
→
Vin bst
vr $ Vin tan b st Vin (73.30 m/s) $ (47.1 m/s) tan (60.0() ! 73.09( 47.1 m/s
The air must now be turned by the rotor blade row in such a way that it leaves the trailing edge of the rotor blade at a zero angle (axially, no swirl) from an absolute reference frame. This determines the rotor’s trailing edge angle, brt. Specifically, when we add an upward velocity of magnitude vr → (Eq. 2) to the relative velocity exiting the trailing edge of the →rotor, V rt, relative, we convert back to the absolute reference frame, →and obtain V rt, the velocity leaving the rotor trailing edge. It is this velocity, V rt, that must be axial (hor→ → izontal). Furthermore, to conserve mass, V rt must equal V in since we are → assuming incompressible flow. Working backward we construct V rt, relative in Fig. 14–66. Trigonometry reveals that
Stator blade row →
Vst
FIGURE 14–64 Velocity vector analysis of the stator blade row of the vane-axial flow fan of Example 14–8; absolute reference frame.
(3)
b rt ! arctan
vr 73.30 m/s ! 57.28( ! arctan Vin 47.1 m/s
(4)
We conclude that the rotor blade at this radius has a leading edge angle of about 73.1° (Eq. 3) and a trailing edge angle of about 57.3° (Eq. 4). A sketch of the rotor blade at this radius is provided in Fig. 14–65; the total curvature is small, being less than 16° from leading to trailing edge. Finally, to avoid interaction of the stator blade wakes with the rotor blade leading edges, we choose the number of rotor blades such that it has no common denominator with the number of stator blades. Since there are 16 stator blades, we pick a number like 13, 15, or 17 rotor blades. Choosing 14 would not be appropriate since it shares a common denominator of 2 with
cen72367_ch14.qxd 11/12/04 2:11 PM Page 773
773 CHAPTER 14
the number 16. Choosing 12 would be worse since it shares both 2 and 4 as common denominators. Discussion We can repeat the calculation for all radii from hub to tip, completing the design of the entire rotor. There would be twist, as discussed previously.
Rotor blade row
→
Vst
vr
14–3
■
PUMP SCALING LAWS
Dimensional Analysis
→
Turbomachinery provides a very practical example of the power and usefulness of dimensional analysis (Chap. 7). We apply the method of repeating variables to the relationship between gravity times net head (gH) and pump . properties such as volume flow rate (V ); some characteristic length, typically the diameter of the impeller blades (D); blade surface roughness height (e); and impeller rotational speed (v), along with fluid properties density (r) and viscosity (m). Note that we treat the group gH as one variable. The dimensionless Pi groups are shown in Fig. 14–67; the result is the following relationship involving dimensionless parameters: # V rvD2 e , b ! function of a 3, m D v 2D2 vD
(14–28)
# V rvD2 e , b ! function of a 3, m D vD rv 3D5
(14–29)
gH
A similar analysis with input brake horsepower as a function of the same variables results in bhp
The second dimensionless parameter (or ) group) on the right side of both Eqs. 14–28 and 14–29 is obviously a Reynolds number since vD is a characteristic velocity, Re !
Vst,relative →
Vrt,relative
brt brl
FIGURE 14–65 Analysis of the stator trailing edge velocity as it impinges on the rotor leading edge; relative reference frame.
vr brt
rvD2 m
→
The third ) on the right is the nondimensional roughness parameter. The three new dimensionless groups in these two equations are given symbols and named as follows:
→
Vrt = Vin →
Vrt, relative
Dimensionless pump parameters: C H ! Head coefficient !
gH v 2D2
# V C Q ! Capacity coefficient ! vD3 C P ! Power coefficient !
(14–30)
bhp rv 3D5
Note the subscript Q in the symbol for capacity coefficient. This comes from the nomenclature found in . many fluid mechanics and turbomachinery textbooks that Q rather than V is the volume flow rate through the pump.
FIGURE 14–66 Analysis of the rotor trailing edge velocity; absolute reference frame.
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774 FLUID MECHANICS
•
V
e D
•
V , r, m
v bhp
C NPSH ! Suction head coefficient !
•
gH = ƒ(V , D, e, v, r, m) k = n – j = 7 – 3 = 4 ∏’s expected. •
gH ∏1 = v2D2 ∏3 =
V ∏2 = vD3
rvD2 m
∏4 =
e D
FIGURE 14–67 Dimensional analysis of a pump.
Pump A
HA •
VA
eA •
DA
VA, rA, mA
bhpA
vA HB
Pump B
•
VB eB •
V B , rB , mB
bhpB
We use the notation C. Q for consistency with turbomachinery convention, even though we use V for volume flow rate to avoid confusion with heat transfer. When pumping liquids, cavitation may be of concern, and we need another dimensionless parameter related to the required net positive suction head. Fortunately, we can simply substitute NPSHrequired in place of H in the dimensional analysis, since they have identical dimensions (length). The result is
DB
vB
FIGURE 14–68 Dimensional analysis is useful for scaling two geometrically similar pumps. If all the dimensionless pump parameters of pump A are equivalent to those of pump B, the two pumps are dynamically similar.
gNPSH required v 2D2
(14–31)
Other variables, such as gap thickness between blade tips and pump housing and blade thickness, can be added to the dimensional analysis if necessary. Fortunately, these variables typically are of only minor importance and are not considered here. In fact, you may argue that two pumps are not even strictly geometrically similar unless gap thickness, blade thickness, and surface roughness scale geometrically. Relationships derived by dimensional analysis, such as Eqs. 14–28 and 14–29, can be interpreted as follows: If two pumps, A and B, are geometrically similar (pump A is geometrically proportional to pump B, although they are of different sizes), and if the independent )’s are equal to each other (in this case if CQ, A ! CQ, B, ReA ! ReB, and eA/DA ! eB/DB), then the dependent )’s are guaranteed to also be equal to each other as well. In particular, CH, A ! CH, B from Eq. 14–28 and CP, A ! CP, B from Eq. 14–29. If such conditions are established, the two pumps are said to be dynamically similar (Fig. 14–68). When dynamic similarity is achieved, the operating point on the pump performance curve of pump A and the corresponding operating point on the pump performance curve of pump B are said to be homologous. The requirement of equality of all three of the independent dimensionless parameters can be relaxed somewhat. If the Reynolds numbers of both pump A and pump B exceed several thousand, turbulent flow conditions exist inside the pump. It turns out that for turbulent flow, if the values of ReA and ReB are not equal, but not too far apart, dynamic similarity between the two pumps is still a reasonable approximation. This fortunate condition is called Reynolds number independence. (Note that if the pumps operate in the laminar regime, the Reynolds number must usually remain as a scaling parameter.) In most cases of practical turbomachinery engineering analysis, the effect of differences in the roughness parameter is also small, unless the roughness differences are large, as when one is scaling from a very small pump to a very large pump (or vice versa). Thus, for many practical problems, we may neglect the effect of both Re and e/D. Equations 14–28 and 14–29 then reduce to C H ! function of C Q
C P ! function of C Q
(14–32)
As always, dimensional analysis cannot predict the shape of the functional relationships of Eq. 14–32, but once these relationships are obtained for a particular pump, they can be generalized for geometrically similar pumps that are of different diameters, operate at different rotational speeds and flow rates, and operate even with fluids of different density and viscosity.
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775 CHAPTER 14
We transform Eq. 14–5 for pump efficiency into a function of the dimensionless parameters of Eq. 14–30,
# 3 2 2 r(V )(gH) r(vD C Q)(v D C H) C QC H h pump ! ! ! function of C Q (14–33) ! bhp CP rv 3D5C P
Since hpump is already dimensionless, it is another dimensionless pump parameter all by itself. Note that since Eq. 14–33 reveals that hpump can be formed by the combination of three other )’s, hpump is not necessary for pump scaling. It is, however, certainly a useful parameter. Since CH, CP, and hpump are functions only of CQ, we often plot these three parameters as functions of CQ on the same plot, generating a set of nondimensional pump performance curves. An example is provided in Fig. 14–69 for the case of a typical centrifugal pump. The curve shapes for other types of pumps would, of course, be different. The simplified similarity laws of Eqs. 14–32 and 14–33 break down when the full-scale prototype is significantly larger than its model (Fig. 14–70); the prototype’s performance is generally better. There are several reasons for this: The prototype pump often operates at high Reynolds numbers that are not achievable in the laboratory. We know from the Moody chart that the friction factor decreases with Re, as does boundary layer thickness. Hence, the influence of viscous boundary layers is less significant as pump size increases, since the boundary layers occupy a less significant percentage of the flow path through the impeller. In addition, the relative roughness (e/D) on the surfaces of the prototype impeller blades may be significantly smaller than that on the model pump blades unless the model surfaces are micropolished. Finally, large full-scale pumps have smaller tip clearances relative to the blade diameter; therefore, tip losses and leakage are less significant. Some empirical equations have been developed to account for the increase in efficiency between a small model and a full-scale prototype. One such equation was suggested by Moody (1926) for turbines, but it can be used as a first-order correction for pumps as well,
BEP
CH
hpump
CH*
CP 0 0
CP* CQ*
CQ
FIGURE 14–69 When plotted in terms of dimensionless pump parameters, the performance curves of all pumps in a family of geometrically similar pumps collapse onto one set of nondimensional pump performance curves. Values at the best efficiency point are indicated by asterisks.
Scale model •
Vmodel
Dmodel
Protoype
Moody efficiency correction equation for pumps: D model 1*5 h pump, prototype ! 1 # (1 # h pump, model)a b D prototype
(14–34)
•
Vprototype
Dprototype
Pump Specific Speed
Another useful dimensionless parameter called pump specific speed (NSp) is formed by a combination of parameters CQ and CH: Pump specific speed:
N Sp !
C Q1/2 CH
! 3/4
# # (V *vD3)1*2 vV 1*2 ! (gH*v 2D2)3*4 (gH)3*4
(14–35)
If all engineers watched their units carefully, NSp would always be listed as a dimensionless parameter. Unfortunately, practicing engineers have grown accustomed to using inconsistent units in Eq. 14–35, which renders the perfectly fine dimensionless parameter NSp into a cumbersome dimensional quantity (Fig. 14–71). Further confusion results because some engineers prefer units of rotations per minute (rpm) for rotational speed, while others
FIGURE 14–70 When a small-scale model is tested to predict the performance of a fullscale prototype pump, the measured efficiency of the model is typically somewhat lower than that of the prototype. Empirical correction equations such as Eq. 14–34 have been developed to account for the improvement of pump efficiency with pump size.
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776 FLUID MECHANICS
You did what? Why would you turn a dimensionless parameter into a dimensional quantity? That’s the exact opposite of what you should be doing!
use rotations per second (Hz), the latter being more common in Europe. In addition, practicing engineers in the United States typically ignore the gravitational constant in the definition of NSp. In this book, we add subscripts “Eur” or “US” to NSp in order to distinguish the dimensional forms of pump specific speed from the nondimensional form. In the United States, it is customary to write H in units of feet (net head. expressed as an equivalent column height of the fluid being pumped), V in units of gallons per minute . (gpm), and rotation rate in terms of n (rpm) instead of v (rad/s). Using Eq. 14–35 we define Pump specific speed, customary U.S. units:
N Sp, US !
# # (n, rpm)(V , gpm)1*2 (H, ft)3*4
(14–36)
In Europe it is customary to write H in units of meters (and to include g . . ! 9.81 m/s2 in the equation), V in units of m3/s, and rotation rate n in units . of rotations per second (Hz) instead of v (rad/s) or n (rpm). Using Eq. 14–35 we define FIGURE 14–71 Even though pump specific speed is a dimensionless parameter, it is common practice to write it as a dimensional quantity using an inconsistent set of units.
Conversion ratios NSp = 3.568 % 10–4 NSp, US NSp NSp, Eur NSp, Eur NSp, US
= 2p
= 5.822 % 10–5
NSp, US NSp
= 2734
NSp, Eur NSp
=
1 2p
NSp,US = 17,180 NSp, Eur
Pump specific speed, customary European units: # # (n, Hz)(V , m3/s)1*2 N Sp, Eur ! (gH, m2/s2)3*4
The conversions between these three forms of pump specific speed are provided as ratios for your convenience in Fig. 14–72. When you become a practicing engineer, you will need to be very careful that you know which form of pump specific speed is being used, although it may not always be obvious. Technically, pump specific speed could be applied at any operating condition and would just be another function of CQ. That is not how it is typically used, however. Instead, it is common to define pump specific speed at only one operating point, namely, the best efficiency point (BEP) of the pump. The result is a single number that characterizes the pump. Pump specific speed is used to characterize the operation of a pump at its optimum conditions (best efficiency point) and is useful for preliminary pump selection.
As plotted in Fig. 14–73, centrifugal pumps perform optimally for NSp near 1, while mixed-flow and axial pumps perform best at NSp near 2 and 5, respectively. It turns out that if NSp is less than about 1.5, a centrifugal pump is the best choice. If NSp is between about 1.5 and 3.5, a mixed-flow pump is a better choice. When NSp is greater than about 3.5, an axial pump should be used. These ranges are indicated in Fig. 14–73 in terms of NSp, NSp, US, and NSp, Eur. Sketches of the blade types are also provided on the plot for reference. EXAMPLE 14–9
FIGURE 14–72 Conversions between the dimensionless, conventional U.S., and conventional European definitions of pump specific speed. Numerical values are given to four significant digits. The conversions for NSp, US assume standard earth gravity.
(14–37)
Using Pump Specific Speed for Preliminary Pump Design
A pump is being designed to deliver 320 gpm of gasoline at room temperature. The required net head is 23.5 ft (of gasoline). It has already been determined that the pump shaft is to rotate at 1170 rpm. Calculate the pump specific speed in both nondimensional form and customary U.S. form. Based on your result, decide which kind of dynamic pump would be most suitable for this application.
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777 CHAPTER 14 NSp, Eur 0.02
1
0.05
0.1
0.2
0.5
Centrifugal
Mixed
1 Axial
0.9 0.8 hmax 0.7 500
0.6
1000
2000
5000
10,000
20,000
NSp, US
0.5 0.1
0.5
0.2
2
1 NSp
5
10
SOLUTION We are to calculate pump specific speed and then determine whether a centrifugal, mixed-flow, or axial pump would be the best choice for this particular application. Assumptions The pump operates near its best efficiency point. 2 The maximum efficiency versus pump specific speed curve follows Fig. 14–73 reasonably well. Analysis First, we calculate pump specific speed in customary U.S. units, N Sp, US !
(1170 rpm)(320 gpm)1$2 (23.5 ft)3$4
! 1960
(1)
We convert to normalized pump specific speed using the conversion factor given in Fig. 14–72,
N Sp b ! 1960(3.658 " 10 #4) ! 0.717 N Sp, US
N Sp ! N Sp, US a
(2)
Using either Eq. 1 or 2, Fig. 14–73 shows that a centrifugal flow pump is the most suitable choice. Discussion Notice that the properties of the fluid never entered our calculations. The fact that we are pumping gasoline rather than some other liquid like water is irrelevant. However, the brake horsepower required to run the pump does depend on the fluid density.
Affinity Laws
We have developed dimensionless groups that are useful for relating any two pumps that are both geometrically similar and dynamically similar. It is convenient to summarize the similarity relationships as ratios. Some authors call these relationships similarity rules, while others call them affinity laws. For any two homologous states A and B,
Affinity laws:
# V B vB D B 3 a b # ! V A vA D A
HB vB 2 D B 2 !a b a b vA HA DA
bhpB r B vB 3 D B 5 ! a b a b bhpA r A vA DA
(14–38a)
(14–38b)
(14–38c)
FIGURE 14–73 Maximum efficiency as a function of pump specific speed for the three main types of dynamic pump. The horizontal scales show nondimensional pump specific speed (NSp), pump specific speed in customary U.S. units (NSp, US), and pump specific speed in customary European units (NSp, Eur).
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778 FLUID MECHANICS ⋅ VB vB 1 n⋅ B 1 ⋅ = av b = a ⋅ b nA VA A
V: Volume flow rate H: Head
HB HA
P: Power
bhpB bhpA
=a =a
2 n⋅ B 2 =a ⋅ b b vA nA
vB
3 n⋅ B 3 b =a ⋅ b vA nA
vB
FIGURE 14–74 When the affinity laws are applied to a single pump in which the only thing that is varied is shaft rotational speed . v, or shaft rpm, n, Eqs. 14–38 reduce to those shown above, for which a jingle can be used to help us remember . the exponent on v (or on n): Very Hard Problems are as easy as 1, 2, 3.
Equations 14–38 apply to both pumps and turbines. States A and B can be any two homologous states between any two geometrically similar turbomachines, or even between two homologous states of the same machine. Examples include changing rotational speed or pumping a different fluid with the same pump. For the simple case of a given pump in which v is varied, but the same fluid is pumped, DA ! DB, and rA ! rB. In such a case, Eqs. 14–38 reduce to the forms shown in Fig. 14–74. A “jingle” has been developed to help us remember the exponent on v, as indicated in the figure. Note also that anywhere there is a ratio of two rotational speeds (v), we . may substitute the appropriate values of rpm (n) instead, since the conversion is the same in both the numerator and the denominator. The pump affinity laws are quite useful as a design tool. In particular, suppose the performance curves of an existing pump are known, and the pump operates with reasonable efficiency and reliability. The pump manufacturer decides to design a new, larger pump for other applications, e.g., to pump a much heavier fluid or to deliver a substantially greater net head. Rather than starting from scratch, engineers often simply scale up an existing design. The pump affinity laws enable such scaling to be accomplished with a minimal amount of effort.
EXAMPLE 14–10
The Effects of Doubling Pump Speed
Professor Seymour Fluids uses a small closed-loop water tunnel to perform flow visualization research. He would like to double the water speed in the test section of the tunnel and realizes that the least expensive way to do this is to double the rotational speed of the flow pump. What he doesn’t realize is how much more powerful the new electric motor will need to be! If Professor Fluids doubles the flow speed, by approximately what factor will the motor power need to be increased? 14
SOLUTION For a doubling of v, we are to calculate by what factor the power to the pump motor must increase. Assumptions 1 The water remains at the same temperature. 2 After doubling pump speed, the pump runs at conditions homologous to the original conditions. Analysis Since neither diameter nor density has changed, Eq. 14–38c reduces to
12 10 8 6
bhpB bhpA
4
Ratio of required shaft power: HB HA
2 0 0
0.5
1
1.5 vB/vA
2
FIGURE 14–75 When the speed of a pump is increased, net head increases very rapidly; brake horsepower increases even more rapidly.
2.5
bhpB vB 3 !a b vA bhpA
(1)
Setting vB ! 2vA in Eq. 1 gives bhpB ! 8bhpA. Thus the power to the pump motor must be increased by a factor of 8. A similar analysis using Eq. 14–38b shows that the pump’s net head increases by a factor of 4. As seen in Fig. 14–75, both net head and power increase rapidly as pump speed is increased. Discussion The result is only approximate since we have not included any analysis of the piping system. While doubling the flow speed through the pump increases available head by a factor of 4, doubling the flow speed through the water tunnel does not necessarily increase the required head of the system by the same factor of 4 (e.g., the friction factor decreases with the Reynolds number except at very high values of Re). In other words, our
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779 CHAPTER 14
assumption 2 is not necessarily correct. The system will, of course, adjust to an operating point at which required and available heads match, but this point will not necessarily be homologous with the original operating point. Nevertheless, the approximation is useful as a first-order result. Professor Fluids may also need to be concerned with the possibility of cavitation at the higher speed.
EXAMPLE 14–11
Design of a New Geometrically Similar Pump
After graduation, you go to work for a pump manufacturing company. One of your company’s best-selling products is a water pump, which we shall call pump A. Its impeller diameter is DA ! 6.0 cm, and its performance data . when operating at nA ! 1725 rpm (vA ! 180.6 rad/s) are shown in Table 14–2. The marketing research department is recommending that the company design a new product, namely, a larger pump (which we shall call pump B) that will be used to pump liquid refrigerant R-134a at room temperature. The pump is to be designed such that its . best efficiency point occurs as close as possible to a volume flow rate of VB ! 2400 cm3/s and at a net head of HB ! 450 cm (of R-134a). The chief engineer (your boss) tells you to perform some preliminary analyses using pump scaling laws to determine if a geometrically scaled-up pump could be designed and built to meet the given requirements. (a) Plot the performance curves of pump A in both dimensional and dimensionless form, and identify the best efficiency point. . (b) Calculate the required pump diameter DB, rotational speed nB, and brake horsepower bhpB for the new product.
TA B L E 1 4 – 2 Manufacturer’s performance data for a water pump operating at 1725 rpm and room temperature (Example 14–11)* . V, cm3/s H, cm hpump, % 100 200 300 400 500 600 700
180 185 175 170 150 95 54
32 54 70 79 81 66 38
* Net head is in centimeters of water.
SOLUTION (a) For a given table of pump performance data for a water pump, we are to plot both dimensional and dimensionless performance curves and identify the BEP. (b) We are to design a new geometrically similar pump for refrigerant R-134a that operates at its BEP at given design conditions. Assumptions 1 The new pump can be manufactured so as to be geometrically similar to the existing pump. 2 Both liquids (water and refrigerant R-134a) are incompressible. 3 Both pumps operate under steady conditions. Properties At room temperature (20°C), the density of water is rwater ! 998.0 kg/m3 and that of refrigerant R-134a is rR-134a ! 1226 kg/m3. Analysis (a) First, we apply a second-order least-squares polynomial curve fit to the data of Table 14–2 to obtain smooth pump performance curves. These are plotted in Fig. 14–76, along with a curve for brake horsepower, which is obtained from Eq. 14–5. A sample. calculation, including unit conversions, is shown in Eq. 1 for the data at VA ! 500 cm3/s, which is approximately the best efficiency point:
# r watergV AHA bhpA ! h pump,A
(998.0 kg/m3)(9.81 m/s2)(500 cm3/s)(150 cm) 1m 4 W's ! a b a b 0.81 100 cm kg ' m/s2 ! 9.07 W
(1)
. Note that the actual value of bhpA plotted in Fig. 14–76 at VA ! 500 cm3/s differs slightly from that of Eq. 1 due to the fact that the least-squares curve fit smoothes out scatter in the original tabulated data.
H, cm (or h, %)
bhp, W
200
10
180
9
bhp
H
160
8
140
7
120
6
100
5
80
4
npump h pump
60
3
40
2
20
1
0
0 0
200
400 600 ⋅ V, cm3/s
800
FIGURE 14–76 Smoothed dimensional pump performance curves for the water pump of Example 14–11.
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780 FLUID MECHANICS
Next we use Eqs. 14–30 to convert the dimensional data of Table 14–2 into nondimensional pump similarity parameters. Sample calculations are shown in Eqs. 2 through 4 at the same . operating point as before (at the approximate location of the BEP). At VA ! 500 cm3/s the capacity coefficient is approximately
# V 500 cm3/s CQ ! ! ! 0.0128 3 vD (180.6 rad/s)(6.0 cm)3
(2)
The head coefficient at this flow rate is approximately
CH !
1.6
CH $ 10 BEP
CP !
1 0.8
hpump
0.6 0.4 0.2 CP $ 100
0 0
0.5
1 CQ $ 100
1.5
FIGURE 14–77 Smoothed nondimensional pump performance curves for the pumps of Example 14–11; BEP is estimated as the operating point where hpump is a maximum.
2
2
vD
!
(9.81 m/s2)(1.50 m) ! 0.125 (180.6 rad/s)2(0.060 m)2
(3)
. Finally, the power coefficient at VA ! 500 cm3/s is approximately
1.4 1.2
gH
2
kg " m/s2 9.07 W ! a b ! 0.00198 (4) rv 3D5 (998 kg/m3)(180.6 rad/s)3(0.060 m)5 W " s bhp
These calculations are repeated (with the aid of a spreadsheet) at values of . VA between 100 and 700 cm3/s. The curve-fitted data are used so that the normalized pump performance curves are smooth; they are plotted in Fig. 14–77. Note that hpump is plotted as a fraction rather than as a percentage. In addition, in order to fit all three curves on one plot with a single ordinate, and with the abscissa centered nearly around unity, we have multiplied CQ by 100, CH by 10, and CP by 100. You will find that these scaling factors work well for a wide range of pumps, from very small to very large. A vertical line at the BEP is also sketched in Fig. 14–77 from the smoothed data. The curve-fitted data yield the following nondimensional pump performance parameters at the BEP:
C Q* ! 0.0112
C H* ! 0.133
C P* ! 0.00184
h*pump ! 0.812
(5)
(b) We design the new pump such that its best efficiency point is homologous with the BEP of the original pump, but with a different fluid, a different pump diameter, and a different rotational speed. Using the values identified in Eq. 5, we use Eqs. 14–30 . to obtain the operating conditions of the new pump. Namely, since both VB and HB are known (design conditions), we solve simultaneously for DB and vB. After some algebra in which we eliminate vB, we calculate the design diameter for pump B,
# 1#4 (0.0024 m3/s)2(0.133) V 2BC H* 1#4 b ! a ! 0.108 m b DB ! a (C *Q)2gHB (0.0112)2(9.81 m/s2)(4.50 m)
(6)
In other words, pump A needs to be scaled up by a factor of DB/DA ! 10.8 cm/6.0 cm ! 1.80. With the value of DB known, we return to Eqs. 14–30 to solve for vB, the design rotational speed for pump B,
# VB 0.0024 m3/s vB ! ! ! 168 rad/s (C *Q)D 3B (0.0112)(0.108 m)3
→
# nB ! 1610 rpm (7)
Finally, the required brake horsepower for pump B is calculated from Eqs. 14–30,
bhpB ! (C*P )r Bv 3BD 5B W"s ! (0.00184)(1226 kg/m3)(168 rad/s)3(0.108 m)5 a b !160 W (8) kg " m2/s
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781 CHAPTER 14
An alternative approach is to use the affinity laws directly, eliminating some intermediate steps. We solve Eqs. 14–38a and b for DB by eliminating the ratio v.B/vA. We then plug in the known value of DA and the curve-fitted values of VA and HA at the BEP (Fig. 14–78). The result agrees with those calculated before. In a similar manner we can calculate vB and bhpB. Discussion Although the desired value of vB has been calculated precisely, a practical issue is that it is difficult (if not impossible) to find an electric motor that rotates at exactly the desired rpm. Standard single-phase, 60-Hz, 120-V AC electric motors typically run at 1725 or 3450 rpm. Thus, we may not be able to meet the rpm requirement with a direct-drive pump. Of course, if the pump is belt-driven or if there is a gear box or a frequency controller, we can easily adjust the configuration to yield the desired rotation rate. Another option is that since vB is only slightly smaller than vA, we drive the new pump at standard motor speed (1725 rpm), providing a somewhat stronger pump than necessary. The disadvantage of this option is that the new pump would then operate at a point not exactly at the BEP.
14–4
■
From the affinity laws,
DB = DA
a
HA HB
= (6.0 cm) a = 10.8 cm
1/4
b a
•
159.3 cm 450 cm
1/2
VB • b VA 1/4
b
2400
cm3 s
438
s
a
1/2
b cm 3
FIGURE 14–78 The affinity laws can be manipulated to obtain an expression for the new pump diameter DB; vB and bhpB can be obtained in similar fashion (not shown).
TURBINES
Turbines have been used for centuries to convert freely available mechanical energy from rivers and wind into useful mechanical work, usually through a rotating shaft. Whereas the rotating part of a pump is called the impeller, the rotating part of a hydroturbine is called the runner. When the working fluid is water, the turbomachines are called hydraulic turbines or hydroturbines. When the working fluid is air, and energy is extracted from the wind, the machine is properly called a wind turbine. The word windmill should technically be applied only when the mechanical energy output is used to grind grain, as in ancient times (Fig. 14–79). However, most people use the word windmill to describe any wind turbine, whether used to grind grain, pump water, or generate electricity. In coal or nuclear power plants, the working fluid is usually steam; hence, the turbomachines that convert energy from the steam into mechanical energy of a rotating shaft are called steam turbines. A more generic name for turbines that employ a compressible gas as the working fluid is gas turbine. (The turbine in a modern commercial jet engine is a type of gas turbine.) In general, energy-producing turbines have somewhat higher overall efficiencies than do energy-absorbing pumps. Large hydroturbines, for example, can achieve overall efficiencies above 95 percent, while the best efficiency of large pumps is a little more than 90 percent. There are several reasons for this. First, pumps normally operate at higher rotational speeds than do turbines; therefore, shear stresses and frictional losses are higher. Second, conversion of kinetic energy into flow energy (pumps) has inherently higher losses than does the reverse (turbines). You can think of it this way: Since pressure rises across a pump (adverse pressure gradient), but drops across a turbine (favorable pressure gradient), boundary layers are less likely to separate in a turbine than in a pump. Third, turbines (especially hydroturbines) are often much larger than pumps, and viscous losses become less important as size increases. Finally, while pumps often operate over a wide range of
FIGURE 14–79 A restored windmill in Brewster, MA, that was used in the 1800s to grind grain. (Note that the blades must be covered to function.) Modern “windmills” that generate electricity are more properly called wind turbines. Courtesy Brewster Historical Society Museum, Brewster, MA. Used by permission.
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782 FLUID MECHANICS
flow rates, most electricity-generating turbines run within a narrower operating range and at a controlled constant speed; they can therefore be designed to operate very efficiently at those conditions. In the United States, the standard AC electrical supply is 60 Hz (3600 cycles per minute); thus most wind, water, and steam turbines operate at speeds that are natural fractions of this, namely, 7200 rpm divided by the number of poles on the generator, usually an even number. Large hydroturbines usually operate at low speeds like 7200/60 ! 120 rpm or 7200/48 ! 150 rpm. Gas turbines used for power generation run at much higher speeds, some up to 7200/2 ! 3600 rpm! As with pumps, we classify turbines into two broad categories, positive displacement and dynamic. For the most part, positive-displacement turbines are small devices used for volume flow rate measurement, while dynamic turbines range from tiny to huge and are used for both flow measurement and power production. We provide details about both of these categories.
Positive-Displacement Turbines
(a)
Shaft
v
Linkage
Flow out
Flow in
(b)
Nutating disc
FIGURE 14–80 The nutating disc fluid flowmeter is a type of positive-displacement turbine used to measure volume flow rate: (a) cutaway view and (b) diagram showing motion of the nutating disc. This type of flowmeter is commonly used as a water meter in homes. Photo courtesy of Niagara Meters, Spartanburg, SC.
A positive-displacement turbine may be thought of as a positive-displacement pump running backward—as fluid pushes into a closed volume, it turns a shaft or displaces a reciprocating rod. The closed volume of fluid is then pushed out as more fluid enters the device. There is a net head loss through the positive-displacement turbine; in other words, energy is extracted from the flowing fluid and is turned into mechanical energy. However, positive-displacement turbines are generally not used for power production, but rather for flow rate or flow volume measurement. The most common example is the water meter in your house (Fig. 14–80). Many commercial water meters use a nutating disc that wobbles and spins as water flows through the meter. The disc has a sphere in its center with appropriate linkages that transfer the eccentric spinning motion of the nutating disc into rotation of a shaft. The volume of fluid that passes through the device per 360o rotation of the shaft is known precisely, and thus the total volume of water used is recorded by the device. When water is flowing at moderate speed from a spigot in your house, you can sometimes hear a bubbly sound coming from the water meter—this is the sound of the nutating disc wobbling inside the meter. There are, of course, other positive-displacement turbine designs, just as there are various designs of positive-displacement pumps.
Dynamic Turbines
Dynamic turbines are used both as flow measuring devices and as power generators. For example, meteorologists use a three-cup anemometer to measure wind speed (Fig. 14–81a). Experimental fluid mechanics researchers use small turbines of various shapes (most of which look like small propellers) to measure air speed or water speed (Chap. 8). In these applications, the shaft power output and the efficiency of the turbine are of little concern. Rather, these instruments are designed such that their rotational speed can be accurately calibrated to the speed of the fluid. Then, by electronically counting the number of blade rotations per second, the speed of the fluid is calculated and displayed by the device. A novel application of a dynamic turbine is shown in Fig. 14–81b. NASA researchers mounted turbines at the wing tips of a Piper PA28
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783 CHAPTER 14
FIGURE 14–81 Examples of dynamic turbines: (a) a typical three-cup anemometer used to measure wind speed, and (b) a Piper PA28 research airplane with turbines designed to extract energy from the wing tip vortices. (a)
(a) Image copyright Campbell Scientific, Inc., Logan, UT, USA. All rights reserved. Used by permission. (b) NASA Langley Research Center.
(b)
research aircraft to extract energy from wing tip vortices (Chap. 11); the extracted energy was converted to electricity to be used for on-board power requirements. In this chapter, we emphasize large dynamic turbines that are designed to produce electricity. Most of our discussion concerns hydroturbines that utilize the large elevation change across a dam to generate electricity. There are two basic types of dynamic turbine—impulse and reaction, each of which are discussed in some detail. Comparing the two power-producing dynamic turbines, impulse turbines require a higher head, but can operate with a smaller volume flow rate. Reaction turbines can operate with much less head, but require a higher volume flow rate.
r v
Impulse Turbines
In an impulse turbine, the fluid is sent through a nozzle so that most of its available mechanical energy is converted into kinetic energy. The highspeed jet then impinges on bucket-shaped vanes that transfer energy to the turbine shaft, as sketched in Fig. 14–82. The modern and most efficient type of impulse turbine was invented by Lester A. Pelton (1829–1908) in 1878, and the rotating wheel is now called a Pelton wheel in his honor. The buckets of a Pelton wheel are designed so as to split the flow in half, and turn the flow nearly 180° around (with respect to a frame of reference moving with the bucket), as illustrated in Fig. 14–82b. According to legend, Pelton modeled the splitter ridge shape after the nostrils of a cow’s nose. A portion of the outermost part of each bucket is cut out so that the majority of the jet can pass through the bucket that is not aligned with the jet (bucket n $ 1 in Fig. 14–82) to reach the most aligned bucket (bucket n in Fig. 14–82). In this way, the maximum amount of momentum from the jet is utilized. These details are seen in a photograph of a Pelton wheel (Fig. 14–83). Figure 14–84 shows a Pelton wheel in operation; the splitting and turning of the water jet is clearly seen. We analyze the power output of a Pelton wheel turbine by using the Euler turbomachine equation. The power output of the shaft is equal to vTshaft, where Tshaft is given by Eq. 14–14, Euler turbomachine equation for a turbine: # # W shaft ! vTshaft ! rvV (r2V2, t # r1V1, t)
(14–39)
Shaft Bucket n + 1 Bucket n
Nozzle
rv
Vj (a)
Splitter ridge
Vj – rv
Vj – rv
b (b)
FIGURE 14–82 Schematic diagram of a Pelton-type impulse turbine; the turbine shaft is turned when high-speed fluid from one or more jets impinges on buckets mounted to the turbine shaft. (a) Side view, absolute reference frame, and (b) bottom view of a cross section of bucket n, rotating reference frame.
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784 FLUID MECHANICS
FIGURE 14–83 A close-up view of a Pelton wheel showing the detailed design of the buckets; the electrical generator is on the right. This Pelton wheel is on display at the Waddamana Power Station Museum near Bothwell, Tasmania. Courtesy of Hydro Tasmania, www.hydro.com.au. Used by permission.
FIGURE 14–84 A view from the bottom of an operating Pelton wheel illustrating the splitting and turning of the water jet in the bucket. The water jet enters from the left, and the Pelton wheel is turning to the right. Courtesy of VA TECH HYDRO. Used by permission.
We must be careful of negative signs since this is an energy-producing rather than an energy-absorbing device. For turbines, it is conventional to define point 2 as the inlet and point 1 as the outlet. The center of the bucket moves at tangential velocity rv, as illustrated in Fig. 14–82. We simplify the analysis by assuming that since there is an opening in the outermost part of each bucket, the entire jet strikes the bucket that happens to be at the direct bottom of the wheel at the instant of time under consideration (bucket n in
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785 CHAPTER 14
Fig. 14–82). Furthermore, since both the size of the bucket and the diameter of the water jet are small compared to the wheel radius, we approximate r1 and r2 as equal to r. Finally, we assume that the water is turned through angle b without losing any speed; in the relative frame of reference moving with the bucket, the relative exit speed is thus Vj # rv (the same as the relative inlet speed) as sketched in Fig. 14–82. Returning to the absolute reference frame, which is necessary for the application of Eq. 14–39, the tangential component of velocity at the inlet, V2, t, is simply the jet speed itself, Vj. We construct a velocity diagram in Fig. 14–85 as an aid in calculating the tangential component of absolute velocity at the outlet, V1, t. After some trigonometry, which you can verify after noting that sin (b # 90°) ! #cos b, V1, t ! rv $ (Vj # rv) cos b
Upon substitution of this equation, Eq. 14–39 yields
# # W shaft ! rrvV {Vj # [rv $ (Vj # rv)cos b]}
which simplifies to Output shaft power:
# # W shaft ! rrvV (Vj # rv)(1 # cos b)
Splitter ridge Vj – rv rv b
Vj –rv
V1,t
→
V1
FIGURE 14–85 Velocity diagram of flow into and out of a Pelton wheel bucket. We translate outflow velocity from the moving reference frame to the absolute reference frame by adding the speed of the bucket (rv) to the right.
(14–40)
Obviously, the maximum power is achieved theoretically if b ! 180°. However, if that were the case, the water exiting one bucket would strike the back side of its neighbor coming along behind it, reducing the generated torque and power. It turns out that in practice, the maximum power is achieved by reducing b to around 160° to 165°. The efficiency factor due to b being less than 180° is Efficiency factor due to b:
# W shaft, actual 1 # cos b hb ! # ! W shaft, ideal 1 # cos (180()
When b ! 160°, for example, hb ! 0.97—a loss of only about. 3 percent. Finally, we see from Eq. 14–40 that . the shaft power output Wshaft is zero if rv ! 0 (wheel not turning at all). Wshaft is also zero if rv ! Vj (bucket moving at the jet speed). Somewhere in between these two extremes lies the optimum wheel speed. By setting the derivative of Eq. 14–40 with respect to rv to zero, we find that this occurs when rv ! Vj /2 (bucket moving at half the jet speed, as shown in Fig. 14–86). For an actual Pelton wheel turbine, there are other losses besides that of Eq. 14–41: mechanical friction, aerodynamic drag on the buckets, friction along the inside walls of the buckets, nonalignment of the jet and bucket as the bucket turns, backsplashing, and nozzle losses. Even so, the efficiency of a well-designed Pelton wheel turbine can approach 90 percent. In other words, up to 90 percent of the available mechanical energy of the water is converted to rotating shaft energy.
Reaction Turbines
Vj v = —– 2r
(14-41)
The other main type of energy-producing hydroturbine is the reaction turbine, which consists of fixed guide vanes called stay vanes, adjustable guide vanes called wicket gates, and rotating blades called runner blades (Fig. 14–87). Flow enters tangentially at high pressure, is turned toward the
r
Shaft Nozzle Vj
Vj rv = —– 2
FIGURE 14–86 The theoretical maximum power achievable by a Pelton turbine occurs when the wheel rotates at v ! Vj /(2r), i.e., when the bucket moves at half the speed of the water jet.
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786 FLUID MECHANICS
Top view Stay vanes
Vin, Pin
Wicket gates
In
Out v
Volute r2
v r1
Runner blades
Shaft
b2 Band
b1 Draft tube
r Out •
Vout, Pout Side view
FIGURE 14–87 A reaction turbine differs significantly from an impulse turbine; instead of using water jets, a volute is filled with swirling water that drives the runner. For hydroturbine applications, the axis is typically vertical. Top and side views are shown, including the fixed stay vanes and adjustable wicket gates.
runner by the stay vanes as it moves along the spiral casing or volute, and then passes through the wicket gates with a large tangential velocity component. Momentum is exchanged between the fluid and the runner as the runner rotates, and there is a large pressure drop. Unlike the impulse turbine, the water completely fills the casing of a reaction turbine. For this reason, a reaction turbine generally produces more power than an impulse turbine of the same diameter, net head, and volume flow rate. The angle of the wicket gates is adjustable so as to control the volume flow rate through the runner. (In most designs the wicket gates can close on each other, cutting off the flow of water into the runner.) At design conditions the flow leaving the wicket gates impinges parallel to the runner blade leading edge (from a rotating frame of reference) to avoid shock losses. Note that in a good design, the number of wicket gates does not share a common denominator with the number of runner blades. Otherwise there would be severe vibration caused by simultaneous impingement of two or more wicket gate wakes onto the leading edges of the runner blades. For example, in Fig. 14–87 there are 17 runner blades and 20 wicket gates. These are typical numbers for many large reaction hydroturbines, as shown in the photographs in Figs. 14–89 and 14–90. The number of stay vanes and wicket gates is usually the same (there are 20 stay vanes in Fig. 14–87). This is not a problem since neither of them rotate, and unsteady wake interaction is not an issue. There are two main types of reaction turbine—Francis and Kaplan. The Francis turbine is somewhat similar in geometry to a centrifugal or mixedflow pump, but with the flow in the opposite direction. Note, however, that a typical pump running backward would not be a very efficient turbine. The Francis turbine is named in honor of James B. Francis (1815–1892), who developed the design in the 1840s. In contrast, the Kaplan turbine is somewhat like an axial-flow fan running backward. If you have ever seen a window fan start spinning in the wrong direction when wind blows hard into the window, you can visualize the basic operating principle of a Kaplan turbine. The Kaplan turbine is named in honor of its inventor, Viktor Kaplan (1876–1934). There are actually several subcategories of both Francis and Kaplan turbines, and the terminology used in the hydroturbine field is not always standard. Recall that we classify dynamic pumps according to the angle at which the flow exits the impeller blade—centrifugal (radial), mixed flow, or axial (see Fig. 14–31). In a similar but reversed manner, we classify reaction turbines according to the angle that the flow enters the runner (Fig. 14–88). If the flow enters the runner radially as in Fig. 14–88a, the turbine is called a Francis radial-flow turbine (see also Fig. 14–87). If the flow enters the runner at some angle between radial and axial (Fig. 14–88b), the turbine is called a Francis mixed-flow turbine. The latter design is more common. Some hydroturbine engineers use the term “Francis turbine” only when there is a band on the runner as in Fig. 14–88b. Francis turbines are most suited for heads that lie between the high heads of Pelton wheel turbines and the low heads of Kaplan turbines. A typical large Francis turbine may have 16 or more runner blades and can achieve a turbine efficiency of 90 to 95 percent. If the runner has no band, and flow enters the runner partially turned, it is called a propeller mixed-flow turbine or simply a mixed-flow
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787 CHAPTER 14 Crown
Crown v
Wicket gate
v
Stay vane Band
Band
(a)
(b) Hub
v
Hub v
(c)
(d)
turbine (Fig. 14–88c). Finally, if the flow is turned completely axially before entering the runner (Fig. 14–88d), the turbine is called an axial-flow turbine. The runners of an axial-flow turbine typically have only three to eight blades, a lot fewer than Francis turbines. Of these there are two types: Kaplan turbines and propeller turbines. Kaplan turbines are called double regulated because the flow rate is controlled in two ways—by turning the wicket gates and by adjusting the pitch on the runner blades. Propeller turbines are nearly identical to Kaplan turbines except that the blades are fixed (pitch is not adjustable), and the flow rate is regulated only by the wicket gates (single regulated). Compared to the Pelton and Francis turbines, Kaplan turbines and propeller turbines are most suited for low head, high volume flow rate conditions. Their efficiencies rival those of Francis turbines and may be as high as 94 percent. Figure 14–89 is a photograph of the radial-flow runner of a Francis radialflow turbine. The workers are shown to give you an idea of how large the runners are in a hydroelectric power plant. Figure 14–90 is a photograph of the mixed-flow runner of a Francis turbine, and Fig. 14–91 is a photograph of the axial-flow propeller of a Kaplan turbine. The view is from the inlet (top). We sketch in Fig. 14–92 a typical hydroelectric dam that utilizes Francis reaction turbines to generate electricity. The overall or gross head Hgross is defined as the elevation difference between the reservoir surface upstream of the dam and the surface of the water exiting the dam, Hgross ! zA # zE. If there were no irreversible losses anywhere in the system, the maximum amount of power that could be generated per turbine would be Ideal power production:
# # W ideal ! rgV Hgross
(14–42)
Of course, there are irreversible losses throughout the system, so the power actually produced is lower than the ideal power given by Eq. 14–42.
FIGURE 14–88 The distinguishing characteristics of the four subcategories of reaction turbines: (a) Francis radial flow, (b) Francis mixed flow, (c) propeller mixed flow, and (d) propeller axial flow. The main difference between (b) and (c) is that Francis mixed-flow runners have a band that rotates with the runner, while propeller mixed-flow runners do not. There are two types of propeller mixed-flow turbines: Kaplan turbines have adjustable pitch blades, while propeller turbines do not. Note that the terminology used here is not universal among turbomachinery textbooks nor among hydroturbine manufacturers.
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FIGURE 14–89 The runner of a Francis radial-flow turbine used at the Round Butte hydroelectric power station in Madras, OR. There are 17 runner blades of outer diameter 11.8 ft (3.60 m). The turbine rotates at 180 rpm and produces 119 MW of power at a volume flow rate of 127 m3/s from a net head of 105 m. Photo courtesy of American Hydro Corporation, York, PA. Used by permission.
FIGURE 14–90 The runner of a Francis mixed-flow turbine used at the Smith Mountain hydroelectric power station in Roanoke, VA. There are 17 runner blades of outer diameter 20.3 ft (6.19 m). The turbine rotates at 100 rpm and produces 194 MW of power at a volume flow rate of 375 m3/s from a net head of 54.9 m. Photo courtesy of American Hydro Corporation, York, PA. Used by permission.
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FIGURE 14–91 The five-bladed propeller of a Kaplan turbine used at the Warwick hydroelectric power station in Cordele, GA. There are five runner blades of outer diameter 12.7 ft (3.87 m). The turbine rotates at 100 rpm and produces 5.37 MW of power at a volume flow rate of 63.7 m3/s from a net head of 9.75 m. Photo courtesy of American Hydro Corporation, York, PA. Used by permission.
A
Gross head Hgross
Dam Head gate (open)
EGL in Power station Generator ⋅ V
Penstock
B
Shaft
C Draft tube
Turbine
D Tailrace
Arbitrary datum plane (z = 0)
Net head H
zA
E
EGLout
zE
FIGURE 14–92 Typical setup and terminology for a hydroelectric plant that utilizes a Francis turbine to generate electricity; drawing not to scale. The Pitot probes are shown for illustrative purposes only.
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790 FLUID MECHANICS
We follow the flow of water through the whole system of Fig. 14–92, defining terms and discussing losses along the way. We start at point A upstream of the dam where the water is still, at atmospheric. pressure, and at its highest elevation, zA. Water flows at volume flow rate V through a large tube through the dam called the penstock. Flow to the penstock can be cut off by closing a large gate valve called a head gate at the penstock inlet. If we were to insert a Pitot probe at point B at the end of the penstock just before the turbine, as illustrated in Fig. 14–92, the water in the tube would rise to a column height equal to the energy grade line EGLin at the inlet of the turbine. This column height is lower than the water level at point A, due to irreversible losses in the penstock and its inlet. The flow then passes through the turbine, which is connected by a shaft to the electric generator. Note that the electric generator itself has irreversible losses. From a fluid mechanics perspective, however, we are interested only in the losses through the turbine and downstream of the turbine. After passing through the turbine runner, the exiting fluid (point C) still has appreciable kinetic energy, and perhaps swirl. To recover some of this kinetic energy (which would otherwise be wasted), the flow enters an expanding area diffuser called a draft tube, which turns the flow horizontally and slows down the flow speed, while increasing the pressure prior to discharge into the downstream water, called the tailrace. If we were to imagine another Pitot probe at point D (the exit of the draft tube), the water in the tube would rise to a column height equal to the energy grade line labeled EGLout in Fig. 14–92. Since the draft tube is considered to be an integral part of the turbine assembly, the net head across the turbine is specified as the difference between EGLin and EGLout, Net head for a hydraulic turbine:
H ! EGLin # EGLout
(14–43)
In words, The net head of a turbine is defined as the difference between the energy grade line just upstream of the turbine and the energy grade line at the exit of the draft tube.
At the draft tube exit (point D) the flow speed is significantly slower than that at point C upstream of the draft tube; however, it is finite. All the kinetic energy leaving the draft tube is dissipated in the tailrace. This represents an irreversible head loss and is the reason why EGLout is higher than the elevation of the tailrace surface, zE. Nevertheless, significant pressure recovery occurs in a well-designed draft tube. The draft tube causes the pressure at the outlet of the runner (point C ) to decrease below atmospheric pressure, thereby enabling the turbine to utilize the available head most efficiently. In other words, the draft tube causes the pressure at the runner outlet to be lower than it would have been without the draft tube—increasing the change in pressure from the inlet to the outlet of the turbine. Designers must be careful, however, because subatmospheric pressures may lead to cavitation, which is undesirable for many reasons, as discussed previously. If we were interested in the net efficiency of the entire hydroelectric plant, we would define this efficiency as the ratio of actual electric power produced to ideal power (Eq. 14–42), based on gross head. Of more concern in
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791 CHAPTER 14
this chapter is the efficiency of the turbine itself. By convention, turbine efficiency is based on net head H rather than gross head Hgross. Specifically, hturbine is defined as the ratio of brake horsepower output (actual turbine output shaft power) to water horsepower (power extracted from the water flowing through the turbine), # bhp W shaft h turbine ! # ! # W water horsepower rgHV
Turbine efficiency:
(14–44)
Note that turbine efficiency hturbine is the reciprocal of pump efficiency hpump, since bhp is the actual output instead of the required input (Fig. 14–93). Note also that we are considering only one turbine at a time in this discussion. Most large hydroelectric power plants have several turbines arranged in parallel. This offers the power company the opportunity to turn off some of the turbines during times of low power demand and for maintenance. Hoover Dam in Boulder City, Nevada, for example, has 17 parallel turbines, 15 of which are identical large Francis turbines that can produce approximately 130 MW of electricity each (Fig. 14–94). The maximum gross head is 590 ft (180 m). The total peak power production of the power plant exceeds 2 GW (2000 MW). We perform preliminary design and analysis of turbines in the same way we did previously for pumps, using the Euler turbomachine equation and velocity diagrams. In fact, we keep the same notation, namely r1 for the inner radius and r2 for the outer radius of the rotating blades. For a turbine, however, the flow direction is opposite to that of a pump, so the inlet is at radius r2 and the outlet is at radius r1. For a first-order analysis we assume that the blades are infinitesimally thin. We also assume that the blades are aligned such that the flow is always tangent to the blade surface, and we
Efficiency is always defined as h = efficiency =
actual output required input
Thus, for a pump, hpump =
⋅ ⋅ Wwater horsepower rgHV = ⋅ bhp Wshaft
and for a turbine, ⋅ Wshaft bhp hturbine = ⋅ = ⋅ Wwater horsepower rgHV
FIGURE 14–93 By definition, efficiency must always be less than unity. The efficiency of a turbine is the reciprocal of the efficiency of a pump.
FIGURE 14–94 (a) An aerial view of Hoover Dam and (b) the top (visible) portion of several of the parallel electric generators driven by hydraulic turbines at Hoover Dam. (a)
(b)
(a) United States Department of the Interior, Bureau of Reclamation— Lower Colorado Region; (b) Photo by Jim Steinhart, PlanetWare.
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792 FLUID MECHANICS
b2
V2, t vr2
→
V2, relative V2, n r1 →
V2
v
r2
FIGURE 14–95 Relative and absolute velocity vectors and geometry for the outer radius of the runner of a Francis turbine. Absolute velocity vectors are bold.
→
b1
V2, relative vr1 →
r1
→
V2
r2
V1
→
V1, relative
v
FIGURE 14–96 Relative and absolute velocity vectors and geometry for the inner radius of the runner of a Francis turbine. Absolute velocity vectors are bold.
ignore viscous effects (boundary layers) at the surfaces. Higher-order corrections are best obtained with a computational fluid dynamics code. Consider for example the top view of the Francis turbine of Fig. 14–87. Velocity vectors are drawn in Fig. 14–95 for both the absolute reference frame and the relative reference frame rotating with the runner. Beginning with the stationary guide vane (thick black line in Fig. 14–95), the flow is turned so that it strikes the runner blade (thick gray line) at absolute veloc→ ity V 2. But the runner blade is rotating counterclockwise, and at radius r2 it moves tangentially to the lower left at speed vr2.→To translate into the rotating reference frame, we form the vector sum of→V 2 and the negative of vr2, as shown in the sketch. The resultant is vector V 2, relative, which is parallel to the runner blade leading edge (angle b2 from the tangent line→of circle r2). The tangential component V2, t, of the absolute velocity vector V 2 is required for the Euler turbomachine equation (Eq. 14–39). After some trigonometry, Runner leading edge:
V2, t ! vr2 #
V2, n tan b 2
(14–45)
Following the flow along the runner blade in the relative (rotating) reference frame, we see that the flow is turned such that it exits parallel to the trailing edge of the runner blade (angle b1 from the tangent line of circle r1). Finally, to translate back to the absolute reference frame we vectorially add → to the left as sketched in Fig. V 1, relative and blade speed vr1, which acts → 14–96. The resultant is absolute vector V 1. Since mass must be conserved, the normal components of the absolute velocity vectors V1, n and V2, n are related through Eq. 14–12, where axial blade widths b1 and b2 are defined in Fig. 14–87. After some trigonometry (which turns out to be identical to that at the leading edge), we generate an →expression for the tangential component V1, t of absolute velocity vector V 1 for use in the Euler turbomachine equation, Runner trailing edge:
V1, t ! vr1 #
V1, n tan b 1
(14–46)
Alert readers will notice that Eq. 14–46 for a turbine is identical to Eq. 14–23 for a pump. This is not just fortuitous, but results from the fact that the velocity vectors, angles, etc., are defined in the same way for a turbine as for a pump except that everything is flowing in the opposite direction. You can see from the Euler turbomachine equation that the maximum power is obtained when V1, t & 0, i.e., when the runner blade turns the flow so much that the swirl at the runner outlet is in the direction opposite to runner rotation. This situation is called reverse swirl (Fig. 14–97). In practice, the fixed-blade runners on most Francis hydroturbines are designed so as to impart a small amount of reverse swirl to the flow exiting the runner. A large amount of reverse swirl is not desirable, however. It turns out that turbine efficiency drops rapidly as the amount of reverse swirl increases because the more the swirl, the more the kinetic energy of the water exiting the turbine, much of which ends up being wasted (draft tubes are not 100 percent efficient). In addition, while reverse swirl may increase output power, the extra turning requires a higher net head for a given volume flow rate. Obviously, much fine-tuning needs to be done in order to design the
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793 CHAPTER 14
most efficient hydroturbine within imposed design constraints. Also keep in mind that the flow is three-dimensional; there is an axial component of velocity as the flow is turned downward into the draft tube. It doesn’t take long before you realize that computer simulation tools are enormously useful to turbine designers. In fact, with the help of modern CFD codes, the efficiency of hydroturbines has increased to the point where retrofits of old turbines in hydroelectric plants are economically wise. An example CFD output is shown in Fig. 14–98 for a Francis mixed-flow turbine. EXAMPLE 14–12
Hydroturbine Design
A retrofit Francis radial-flow hydroturbine is being designed to replace an old turbine in a hydroelectric dam. The new turbine must meet the following design restrictions in order to properly couple with the existing setup: The runner inlet radius is r2 ! 8.20 ft (2.50 m) and its outlet radius is r1 ! 5.80 ft (1.77 m). The runner blade widths are b2 ! 3.00 ft (0.914 m) and b1 ! 8.60 ft (2.62 m) at the inlet and outlet, respectively. The runner . must rotate at n ! 120 rpm (v ! 12.57 rad/s) to turn the 60-Hz electric generator. The wicket gates turn the flow by angle a2 ! 33° from radial at the runner inlet, and the flow at the runner outlet is to have angle a1
6.00e + 05 5.70e + 05 5.40e + 05 5.10e + 05 4.80e + 05 4.50e + 05 4.20e + 05 3.90e + 05 3.60e + 05 3.30e + 05 3.00e + 05 2.70e + 05 2.40e + 05 2.10e + 05 1.80e + 05 1.50e + 05 1.20e + 05 9.00e + 04 6.00e + 04 3.00e + 04 0.00e + 00 –3.00e + 04 –6.00e + 04 –9.00e + 04 –1.20e + 05 –1.50e + 05
z x y
v
Reverse swirl
FIGURE 14–97 In many modern Francis mixed-flow hydroturbines, the flow exiting the runner swirls in a direction opposite to that of the runner itself. This is called reverse swirl.
FIGURE 14–98 Static pressure distribution (grayscale contour plot) on runner blade surfaces as calculated by CFD; pressure is in units of pascals. Shown is a 17-blade Francis mixed-flow turbine runner that rotates counterclockwise about the z-axis. Only one blade passage is modeled, but the image is reproduced 16 times due to the symmetry. The highest pressures (light regions) are encountered near the leading edges of the pressure surfaces of the runner, while the lowest pressures (dark regions) occur on the suction surface of the runner near the trailing edge. Photo courtesy of American Hydro Corporation, York, PA. Used by permission.
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794 FLUID MECHANICS
V2,
t
a2
between #10° and 10° from radial (Fig. 14–99) for proper flow through the draft tube. The volume flow rate at design conditions is 9.50 % 106 gpm (599 m3/s), and the gross head provided by the dam is Hgross ! 303 ft (92.4 m). (a) Calculate the inlet and outlet runner blade angles b2 and b1, respectively, and predict the power output and required net head if irreversible losses are neglected for the case with a1 ! 10° from radial (withrotation swirl). (b) Repeat the calculations for the case with a1 ! 0° from radial (no swirl). (c) Repeat the calculations for the case with a1 ! #10° from radial (reverse swirl).
V2 n
→
V2 V1, t
r2
→
a1
V1
V1, n
r1 v
Control volume
FIGURE 14–99 Top view of the absolute velocities and flow angles associated with the runner of a Francis turbine being designed for a hydroelectric dam. The control volume is from the inlet to the outlet of the runner.
SOLUTION For a given set of hydroturbine design criteria we are to calculate runner blade angles, required net head, and power output for three cases—two with swirl and one without swirl at the runner outlet. Assumptions 1 The flow is steady. 2 The fluid is water at 20°C. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties For water at 20°C, r ! 998.0 kg/m3. Analysis (a) We solve for the normal component of velocity at the inlet using Eq. 14-12, # V 599 m3/s V2, n ! ! ! 41.7 m/s 2pr2b2 2p(2.50 m)(0.914 m)
(1)
Using Fig. 14–99 as a guide, the tangential velocity component at the inlet is
V2, t ! V2, n tan a 2 ! (41.7 m/s) tan 33( ! 27.1 m/s
(2)
We now solve Eq. 14-45 for the runner leading edge angle b2,
V2, n 41.7 m/s b 2 ! arctan a b ! arctana b ! 84.1! (3) vr2 # V2, t (12.57 rad/s)(2.50 m) # 27.1 m/s
a2 →
V2
Equations 1 through 3 are repeated for the runner outlet, with the following results:
b2
Runner outlet:
V1, t ! 3.63 m/s,
b 1 ! 47.9!
(4)
The top view of this runner blade is sketched (to scale) in Fig. 14–100. Using Eqs. 2 and 4, the shaft output power is estimated from the Euler turbomachine equation, Eq. 14–39,
b1 →
# # Wshaft ! rvV(r2V2, t # r1V1, t) ! (998.0 kg/m 3)(12.57 rads/s)(599 m3/s)
V1
r2
V1, n ! 20.6 m/s,
a1
% [(2.50 m)(27.2 m/s) # (1.77 m)(3.63 m/s)]a
r1
! 461 MW ! 6.18 % 105 hp
v
FIGURE 14–100 Sketch of the runner blade design of Example 14-12, top view. A guide vane and absolute velocity vectors are also shown.
MW ' s b 10 6 kg ' m2/s2
(5)
Finally, we calculate the required net head using Eq. 14–44, assuming that hturbine ! 100 percent since we are ignoring irreversibilities,
H!
10 6 kg ' m2/s2 bhp 461 MW b !78.6 m (6) a #! 3 2 3 MW ' s rgV (998.0 kg/m )(9.81 m/s )(599 m /s)
(b) When we repeat the calculations with no swirl at the runner outlet (a1 ! 0°), the runner blade trailing edge angle reduces to 42.8°, and the output
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795 CHAPTER 14
power increases to 509 MW (6.83 " 105 hp). The required net head increases to 86.8 m. (c) When we repeat the calculations with reverse swirl at the runner outlet (a1 ! #10°), the runner blade trailing edge angle reduces to 38.5°, and the output power increases to 557 MW (7.47 " 105 hp). The required net head increases to 95.0 m. A plot of power and net head as a function of runner outlet flow angle a1 is shown in Fig. 14–101. You can see that both bhp and H increase with decreasing a1. Discussion The theoretical output power increases by about 10 percent by eliminating swirl from the runner outlet and by nearly another 10 percent when there is 10° of reverse swirl. However, the gross head available from the dam is only 92.4 m. Thus, the reverse swirl case of part (c) is clearly impossible, since the predicted net head is required to be greater than Hgross. Keep in mind that this is a preliminary design in which we are neglecting irreversibilities. The actual output power will be lower and the actual required net head will be higher than the values predicted here.
Hgross
100
■
TURBINE SCALING LAWS
Dimensionless Turbine Parameters
We define dimensionless groups (Pi groups) for turbines in much the same way as we did in Section 14–3 for pumps. Neglecting Reynolds number and roughness effects, we deal with the same . dimensional variables: gravity times net head (gH), volume flow rate (V ), diameter of the runner blades (D), runner rotational speed (v), output brake horsepower (bhp), and fluid density (r), as illustrated in Fig. 14–102. In fact, the dimensional analysis is identical whether analyzing a pump . or a turbine, except for the fact that for turbines, we take bhp instead of V as the independent variable. In addition, hturbine (Eq. 14–44) is used in place of hpump as the nondimensional efficiency. A summary of the dimensionless parameters is provided here: Dimensionless turbine parameters: C H ! Head coefficient !
v 2D2
500 bhp
60
400 300
40
200 20
100 0
0 –10
0 10 a1, degrees
20
FIGURE 14–101 Ideal required net head and brake horsepower output as functions of runner outlet flow angle for the turbine of Example 14–12.
(14–47)
bhp rv 3D5
h turbine ! Turbine efficiency !
CP ! function of C P C QC H
bhp
bhp
H = net head
# rgHV
When plotting turbine performance curves, we use CP instead of CQ as the independent parameter. In other words, CH and CQ are functions of CP, and hturbine is thus also a function of CP, since h turbine !
600
# V C Q ! Capacity coefficient ! vD3
gH
C P ! Power coefficient !
700
H
80
–20
14–5
bhp, MW
H, m
v
D
(14–48)
The affinity laws (Eqs. 14–38) can be applied to turbines as well as to pumps, allowing us to scale turbines up or down in size (Fig. 14–103). We also use the affinity laws to predict the performance of a given turbine operating at different speeds and flow rates in the same way as we did previously for pumps.
r •
V
FIGURE 14–102 The main variables used for dimensional analysis of a turbine.
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796 FLUID MECHANICS Turbine A
bhpA HA = net head vA
DA
rA
•
VA bhpB
Turbine B
HB = net head vB
DB
The simple similarity laws are strictly valid only if the model and the prototype operate at identical Reynolds numbers and are exactly geometrically similar (including relative surface roughness and tip clearance). Unfortunately, it is not always possible to satisfy all these criteria when performing model tests, because the Reynolds number achievable in the model tests is generally much smaller than that of the prototype, and the model surfaces have larger relative roughness and tip clearances. When the full-scale prototype is significantly larger than its model, the prototype’s performance is generally better, for the same reasons discussed previously for pumps. Some empirical equations have been developed to account for the increase in efficiency between a small model and a full-scale prototype. One such equation was suggested by Moody (1926), and can be used as a first-order correction, Moody efficiency correction equation for turbines: D model 1*5 h turbine, prototype ! 1 # (1 # h turbine, model)a b D prototype
rB
•
VB
FIGURE 14–103 Dimensional analysis is useful for scaling two geometrically similar turbines. If all the dimensionless turbine parameters of turbine A are equivalent to those of turbine B, the two turbines are dynamically similar.
(14–49)
Note that Eq. 14–49 is also used as a first-order correction when scaling model pumps to full scale (Eq. 14–34). In practice, hydroturbine engineers generally find that the actual increase in efficiency from model to prototype is only about two-thirds of the increase given by Eq. 14–49. For example, suppose the efficiency of a onetenth scale model is 93.2 percent. Equation 14–49 predicts a full-scale efficiency of 95.7 percent, or an increase of 2.5 percent. In practice, we expect only about two-thirds of this increase, or 93.2 $ 2.5(2/3) ! 94.9 percent. Some more advanced correction equations are available from the International Electrotechnical Commission (IEC), a worldwide organization for standardization. EXAMPLE 14–13
Application of Turbine Affinity Laws
A Francis turbine is being designed for a hydroelectric dam. Instead of starting from scratch, the engineers decide to geometrically scale up a previously designed hydroturbine that has an excellent performance history. The existing . turbine (turbine A) has diameter DA ! 2.05 m, .and spins at nA ! 120 rpm 3 (vA ! 12.57 rad/s). At its best efficiency point, VA ! 350 m /s, HA ! 75.0 m of water, and bhpA ! 242 MW. The new turbine (turbine B) is for a larger facility. Its generator will spin at the same speed (120 rpm), but its net head will be higher (HB ! 104 m). Calculate the diameter of the new turbine such . that it operates most efficiently, and calculate VB, bhpB, and hturbine, B.
SOLUTION We are to design a new hydroturbine by scaling up an existing hydroturbine. Specifically we are to calculate the new turbine diameter, volume flow rate, and brake horsepower. Assumptions 1 The new turbine is geometrically similar to the existing turbine. 2 Reynolds number effects and roughness effects are negligible. 3 The new penstock is also geometrically similar to the existing penstock so that the flow entering the new turbine (velocity profile, turbulence intensity, etc.) is similar to that of the existing turbine. Properties The density of water at 20°C is r ! 998.0 kg/m3.
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797 CHAPTER 14
Analysis Since the new turbine (B) is dynamically similar to the existing turbine (A), we are concerned with only one particular homologous operating point of both turbines, namely, the best efficiency point. We solve Eq. 14–38b for DB,
# HB nA 104 m 120 rpm ! 2.41 m DB ! DA # ! (2.05m) B HA nB B 75.0 m 120 rpm
. We then solve Eq. 14–38a for VB,
# # # nB D B 3 120 rpm 2.41 m 3 V B ! V A a # b a b ! (350 m3/s) a ba b ! 572 m3/s nA D A 120 rpm 2.05 m
Finally, we solve Eq. 14–38c for bhpB,
# r B nB 3 D B 5 ba# b a b r A nA DA
bhpB ! bhpA a
CH, A = CH, B =
998.0 kg/m3
120 rpm 3 2.41 m 5 ! (242 MW)a b a b ! 548 MW b a 2.05 m 998.0 kg/m3 120 rpm
CQ, A = CQ, B =
As a check, we calculate the dimensionless turbine parameters of Eq. 14–47 for both turbines to show that these two operating points are indeed homologous (Fig. 14–104). As discussed previously, however, total dynamic similarity may not actually be achieved between the two turbines because of scale effects (larger turbines generally have higher efficiency). The diameter of the new turbine is about 18 percent greater than that of the existing turbine, so the increase in efficiency due to turbine size should not be significant. We verify this by using the Moody efficiency correction equation (Eq. 14–49), considering turbine A as the “model” and B as the “prototype,”
Efficiency correction: D A 1*5 2.05 m 1*5 h turbine, B ! 1 # (1 # h turbine, A)a b !1 # (1 # 0.942)a b ! 0.944 DB 2.41 m
or 94.4 percent. Indeed, the first-order correction yields a predicted efficiency for the larger turbine that is only a fraction of a percent greater than that of the smaller turbine. Discussion If the flow entering the new turbine from the penstock were not similar to that of the existing turbine (e.g., velocity profile and turbulence intensity), we could not expect exact dynamic similarity.
Turbine Specific Speed
In our previous discussion of pumps, we defined another useful dimensionless parameter, pump specific speed (NSp), based on CQ and CH. We could use the same definition of specific speed for turbines, but since CP rather than CQ is the independent dimensionless parameter for turbines, we define turbine specific speed (NSt) differently, namely, in terms of CP and CH, Turbine specific speed: N St !
C 1/2 P
(bhp/rv 3D5)1*2
CH
(gH/v 2D2)5*4
! 5/4
!
v(bhp)1*2 r 1*2(gH)5*4
(14–50)
CP, A = CP, B =
gH v2D2
= 1.11
⋅ V vD3
= 3.23
bhp rv3D5
hturbine, A = hturbine, B =
= 3.38 bhp ⋅ = 94.2% rgHV
FIGURE 14–104 Dimensionless turbine parameters for both turbines of Example 14–13. Since the two turbines operate at homologous points, their dimensionless parameters must match.
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798 FLUID MECHANICS
Motor/generator (acting as a motor)
Turbine specific speed is also called power specific speed in some textbooks. It is left as an exercise to compare the definitions of pump specific speed (Eq. 14–35) and turbine specific speed (Eq. 14–50) in order to show that Relationship between NSt and NSp:
Pump–turbine (acting as a pump) (a)
Motor/generator (acting as a generator)
Pump–turbine (acting as a turbine) (b)
FIGURE 14–105 A pump–turbine is used by some power plants for energy storage: (a) water is pumped by the pump– turbine during periods of low demand for power, and (b) electricity is generated by the pump–turbine during periods of high demand for power.
N St ! N Sp 2h turbine
(14–51)
Note that Eq. 14–51 does not apply to a pump running backward as a turbine or vice versa. There are applications in which the same turbomachine is used as both a pump and a turbine; these devices are appropriately called pump–turbines. For example, a coal or nuclear power plant may pump water to a higher elevation during times of low power demand, and then run that water through the same turbomachine (operating as a turbine) during times of high power demand (Fig. 14–105). Such facilities often take advantage of natural elevation differences at mountainous sites and can achieve significant gross heads (upward of 1000 ft) without construction of a dam. A photograph of a pump–turbine is shown in Fig. 14–106. Note that there are inefficiencies in the pump–turbine when operating as a pump and also when operating as a turbine. Moreover, since one turbomachine must be designed to operate as both a pump and a turbine, neither hpump nor hturbine are as high as they would be for a dedicated pump or turbine. Nevertheless, the overall efficiency of this type of energy storage is around 80 percent for a well-designed pump–turbine unit. In practice, the pump–turbine may operate at a different flow rate and rpm when it is acting as a turbine compared to when it is acting as a pump, since the best efficiency point of the turbine is not necessarily the same as that of the pump. However, for the simple case in which the flow rate and rpm are the same for both the pump and turbine operations, we use Eqs. 14–35 and 14–50 to compare pump specific speed and turbine specific speed. After some algebra, Pump–turbine specific speed relationship at same flow rate and rpm: Hpump 3*4 bhppump 3*4 N St ! N Sp 2h turbine a b ! N Sp(h turbine)5*4(h pump)3*4 a b Hturbine bhpturbine
FIGURE 14–106 The runner of a pump–turbine used at the Yards Creek pumped storage station in Blairstown, NJ. There are seven runner blades of outer diameter 17.3 ft (5.27 m). The turbine rotates at 240 rpm and produces 112 MW of power at a volume flow rate of 56.6 m3/s from a net head of 221 m. Photo courtesy of American Hydro Corporation, York, PA. Used by permission.
(14–52)
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799 CHAPTER 14
We previously discussed some problems with the units of pump specific speed. Unfortunately, these same problems also occur with turbine specific speed. Namely, although NSt is by definition a dimensionless parameter, practicing engineers have grown accustomed to using inconsistent units that transform NSt into a cumbersome dimensional quantity. In the United States, most turbine engineers write the rotational speed in units of rotations per minute (rpm), bhp in units of horsepower, and H in units of feet. Furthermore, they ignore gravitational constant g and density r in the definition of NSt. (The turbine is assumed to operate on earth and the working fluid is assumed to be water.) We define Turbine specific speed, customary U.S. units: # (n, rpm) (bhp, hp)1*2 N St, US ! (H, ft)5*4
Conversion ratios
(14–53)
There is some discrepancy in the turbomachinery literature over the conversions between the two forms of turbine specific speed. To convert NSt, US to NSt we divide by g5/4 and r1/2, and then use conversion ratios to cancel all units. We set g ! 32.174 ft/s2 and assume water at density r ! 62.40 lbm/ft3. When done properly by converting v to rad/s, the conversion is NSt, US ! 0.02301NSt or NSt ! 43.46NSt, US. However, some authors convert v to rotations per second, introducing a factor of 2p in the conversion, i.e., NSt, US ! 0.003662NSt or NSt ! 273.1NSt, US. The former conversion is more common and is summarized in Fig. 14–107. Technically, turbine specific speed could be applied at any operating condition and would just be another function of CP. That is not how it is typically used, however. Instead, it is common to define turbine specific speed only at the best efficiency point (BEP) of the turbine. The result is a single number that characterizes the turbine.
NSt = 0.02301 NSt, US NSt, US = 43.46 NSt
FIGURE 14–107 Conversions between the dimensionless and the conventional U.S. definitions of turbine specific speed. Numerical values are given to four significant digits. The conversions assume earth gravity and water as the working fluid.
Turbine specific speed is used to characterize the operation of a turbine at its optimum conditions (best efficiency point) and is useful for preliminary turbine selection.
As plotted in Fig. 14–108, impulse turbines perform optimally for NSt near 0.15, while Francis turbines and Kaplan turbines perform best at NSt near 1 and 2.5, respectively. It turns out that if NSt is less than about 0.3, an
NSt, US 1
1
2
5
10
20
50
100
Francis
200
Kaplan
Impulse
0.9 0.8 hmax 0.7 0.6 0.5 0.01
0.02
0.05
0.1
0.2
0.5 NSt
1
2
5
10
FIGURE 14–108 Maximum efficiency as a function of turbine specific speed for the three main types of dynamic turbine. Horizontal scales show nondimensional turbine specific speed (NSt) and turbine specific speed in customary U.S. units (NSt, US). Sketches of the blade types are also provided on the plot for reference.
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impulse turbine is the best choice. If NSt is between about 0.3 and 2, a Francis turbine is a better choice. When NSt is greater than about 2, a Kaplan turbine should be used. These ranges are indicated in Fig. 14–108 in terms of NSt and NSt, US. EXAMPLE 14–14
Turbine Specific Speed
Calculate and compare the turbine specific speed for both the small (A) and large (B) turbines of Example 14–13.
SOLUTION The turbine specific speed of two dynamically similar turbines is to be compared. Properties The density of water at T ! 20°C is r ! 998.0 kg/m3. Analysis We calculate the dimensionless turbine specific speed for turbine A, N St, A ! !
vA(bhpA)1*2 5*4 r 1*2 A (gHA)
kg ' m/s2 1*2 (12.57 rad/s)(242 % 10 6 W)1*2 a b ! 1.615 ! 1.62 (998.0 kg/m3)1*2[(9.81 m/s2)(75.0 m)]5*4 W ' s
and for turbine B, Turbine Specific Speed: NSt =
CP1/2 CH5/4
=
(3.38)1/2 (1.11)5/4
N St, B ! = 1.61
FIGURE 14–109 Calculation of turbine specific speed using the dimensionless parameters CP and CH for Example 14–14. (See Fig. 14–104 for values of CP and CH for turbine A and turbine B.)
!
vB(bhpB)1*2 5*4 r 1*2 B (gHB)
kg ' m/s2 1*2 (12.57 rad/s)(548 % 10 6 W)1*2 a b ! 1.615 ! 1.62 (998.0 kg/m3)1*2 [(9.81 m/s2)(104 m)]5*4 W ' s
We see that the turbine specific speeds of the two turbines are the same. As a check of our algebra we calculate NSt in Fig. 14–109 a different way using its definition in terms of CP and CH (Eq. 14–50). The result is the same (except for roundoff error). Finally, we calculate the turbine specific speed in customary U.S. units from the conversions of Fig. 14–107,
N St, US, A ! N St, US, B ! 43.46N St ! (43.46)(1.615) ! 70.2 Discussion Since turbines A and B operate at homologous points, it is no surprise that their turbine specific speeds are the same. In fact, if they weren’t the same, it would be a sure sign of an algebraic or calculation error. From Fig. 14–108, a Francis turbine is indeed the appropriate choice for a turbine specific speed of 1.6.
Gas and Steam Turbines
Most of our discussion so far has concerned hydroturbines. We now discuss turbines that are designed for use with gases, like combustion products or steam. In a coal or nuclear power plant, high-pressure steam is produced by a boiler and then sent to a steam turbine to produce electricity. Because of reheat, regeneration, and other efforts to increase overall efficiency, these steam turbines typically have two stages (high pressure and low pressure). Most power plant steam turbines are multistage axial-flow devices like that shown in Fig. 14–110. Not shown are the stator vanes that direct the flow between each set of turbine blades. Analysis of axial-flow turbines is very similar to that of axial-flow fans, as discussed in Section 14–2, and will not be repeated here.
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Similar axial-flow turbines are used in jet aircraft engines (Fig. 14–62) and gas turbine generators (Fig. 14–111). A gas turbine generator is similar to a jet engine except that instead of providing thrust, the turbomachine is designed to transfer as much of the fuel’s energy as possible into the rotating shaft, which is connected to an electric generator. Gas turbines used for power generation are typically much larger than jet engines, of course, since they are ground-based. As with hydroturbines, a significant gain in efficiency is realized as overall turbine size increases.
FIGURE 14–110 The turbine blades of a typical twostage steam turbine used in a coal or nuclear power plant. The flow is from left to right, with the high-pressure stage on the left and the low-pressure stage on the right. Courtesy, Alstom. Used by permission.
FIGURE 14–111 The rotor assembly of the MS7001F gas turbine being lowered into the bottom half of the gas turbine casing. Flow is from right to left, with the upstream set of rotor blades (called blades) comprising the multistage compressor and the downstream set of rotor blades (called buckets) comprising the multistage turbine. Compressor stator blades (called vanes) and turbine stator blades (called nozzles) can be seen in the bottom half of the gas turbine casing. This gas turbine spins at 3600 rpm and produces over 135 MW of power. Courtesy General Electric Power Systems. Used by permission.
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APPLICATION SPOTLIGHT
■
Rotary Fuel Atomizers
Guest Author: Werner J. A. Dahm, The University of Michigan d
t t R2
Fuel: rL, mL
R1
v
d (a)
(b)
FIGURE 14–112 Schematic diagram of (a) a rotary fuel atomizer, and (b) a close-up of the liquid fuel film along the channel walls.
FIGURE 14–113 Visualizations of liquid breakup by rotary fuel atomizers, showing subcritical breakup at relatively low values of Wet (top), for which surface tension effects are sufficiently strong relative to inertia to pull the thin liquid film into large columns; and supercritical breakup at higher values of Wet (bottom), for which inertia dominates over surface tension and the thin film breaks into fine droplets. Reprinted by permission of Werner J. A. Dahm, University of Michigan.
The very high rotation rates at which small gas turbine engines operate, often approaching 100,000 rpm, allow rotary centrifugal atomizers to create the liquid fuel spray that is burned in the combustor. Note that a 10-cm-diameter atomizer rotating at 30,000 rpm imparts 490,000 m/s2 of acceleration (50,000 g) to the liquid fuel, which allows such fuel atomizers to potentially produce very small drop sizes. The actual drop sizes depend on the fluid properties, including the liquid and gas densities rL and rG, the viscosities mL and mG, and the liquid–gas surface tension ss. Figure 14–112 shows such a rotary atomizer rotating at rate v, with radial channels in the rim at nominal radius R # (R1 $ R2)/2. Fuel flows into the channels due to the acceleration Rv2 and forms a liquid film on the channel walls. The large acceleration leads to a typical film thickness t of only about 10 mm. The channel shape is chosen to produce desirable atomization performance. For a given shape, the resulting drop sizes depend on the cross-flow velocity Vc # Rv into which the film issues at the channel exit, together with the liquid and gas properties. From these, there are four dimensionless groups that determine the atomization performance: the liquid–gas density and viscosity ratios r # [rL/rG] and m # [mL/mG], the film Weber number Wet # [rGV c2t/ss], and the Ohnesorge number Oht # [mL/(rLsst)1/2]. Note that Wet gives the characteristic ratio of the aerodynamic forces that the gas exerts on the liquid film to the surface tension forces that act on the liquid surface, while Oht gives the ratio of the viscous forces in the liquid film to the surface tension forces that act on the film. Together these express the relative importance of the three main physical effects involved in the atomization process: inertia, viscous diffusion, and surface tension. Figure 14–113 shows examples of the resulting liquid breakup process for several channel shapes and rotation rates, visualized using 10-ns pulsed-laser photography. The drop sizes turn out to be relatively insensitive to changes in the Ohnesorge number, since the values for practical fuel atomizers are in the limit Oht && 1 and thus viscous effects are relatively unimportant. The Weber number, however, remains crucial since surface tension and inertia effects dominate the atomization process. At small Wet, the liquid undergoes subcritical breakup in which surface tension pulls the thin liquid film into a single column that subsequently breaks up to form relatively large drops. At supercritical values of Wet, the thin liquid film breaks up aerodynamically into fine drop sizes on the order of the film thickness t. From results such as these, engineers can successfully develop rotary fuel atomizers for practical applications. Reference Dahm, W. J. A., Patel, P. R., and Lerg, B. H., “Visualization and Fundamental Analysis of Liquid Atomization by Fuel Slingers in Small Gas Turbines,” AIAA Paper No. 2002-3183, AIAA, Washington, DC, 2002.
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SUMMARY We classify turbomachinery into two broad categories, pumps and turbines. The word pump is a general term for any fluid machine that adds energy to a fluid. We explain how this energy transfer occurs for several types of pump designs— both positive-displacement pumps and dynamic pumps. The word turbine refers to a fluid machine that extracts energy from a fluid. There are also positive-displacement turbines and dynamic turbines of several varieties. The most useful equation for preliminary turbomachinery design is the Euler turbomachine equation, # Tshaft ! rV (r2V2, t # r1V1, t) Note that for pumps, the inlet and outlet are at radii r1 and r2, respectively, while for turbines, the inlet is at radius r2 and the outlet is at radius r1. We show several examples where blade shapes for both pumps and turbines are designed based on desired flow velocities. Then, using the Euler turbomachine equation, the performance of the turbomachine is predicted. The turbomachinery scaling laws illustrate a practical application of dimensional analysis. The scaling laws are used in the design of new turbomachines that are geometrically similar to existing turbomachines. For both pumps and turbines, the main dimensionless parameters are head coefficient, capacity coefficient, and power coefficient, defined respectively as # bhp gH V CQ ! CP ! 3 5 CH ! 2 2 vD vD3 rv D
In addition to these, we define pump efficiency and turbine efficiency as reciprocals of each other, # # W water horsepower rgV H ! h pump ! # bhp W shaft # bhp W shaft h turbine ! # ! # W water horsepower rgV H Finally, two other useful dimensionless parameters called pump specific speed and turbine specific speed are defined, respectively, as # C1/2 v(bhp)1*2 C1/2 vV 1*2 Q P N ! ! N Sp ! 3/4 ! St CH (gH)3*4 C 5/4 r 1*2(gH)5*4 H These parameters are useful for preliminary selection of the type of pump or turbine that is most appropriate for a given application. Turbomachinery design assimilates knowledge from several key areas of fluid mechanics, including mass, energy, and momentum analysis (Chaps. 5 and 6); dimensional analysis and modeling (Chap. 7); flow in pipes (Chap. 8); differential analysis (Chaps. 9 and 10); and aerodynamics (Chap. 11). In addition, for gas turbines and other types of turbomachines that involve gases, compressible flow analysis (Chap. 12) is required. Finally, computational fluid dynamics (Chap. 15) plays an ever-increasing role in the design of highly efficient turbomachines.
REFERENCES AND SUGGESTED READING 1. ASHRAE (American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc.). ASHRAE Fundamentals Handbook, ASHRAE, 1791 Tullie Circle, NE, Atlanta, GA, 30329; editions every four years: 1993, 1997, 2001, etc. 2. L. F. Moody. “The Propeller Type Turbine,” ASCE Trans., 89, p. 628, 1926. 3. Earl Logan, Jr., ed. Handbook of Turbomachinery. New York: Marcel Dekker, Inc., 1995. 4. A. J. Glassman, ed. Turbine Design and Application. NASA Sp-290, NASA Scientific and Technical Information Program. Washington, DC, 1994.
5. D. Japikse and N. C. Baines. Introduction to Turbomachinery. Norwich, VT: Concepts ETI, Inc., and Oxford: Oxford University Press, 1994. 6. Earl Logan, Jr. Turbomachinery: Basic Theory and Applications, 2nd ed. New York: Marcel Dekker, Inc., 1993. 7. R. K. Turton. Principles of Turbomachinery, 2nd ed. London: Chapman & Hall, 1995. 8. Terry Wright. Fluid Machinery: Performance, Analysis, and Design. Boca Raton, FL: CRC Press, 1999.
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PROBLEMS* General Problems 14–1C What is the more common term for an energyproducing turbomachine? How about an energy-absorbing turbomachine? Explain this terminology. In particular, from which frame of reference are these terms defined—that of the fluid or that of the surroundings? 14–2C What are the primary differences between fans, blowers, and compressors? Discuss in terms of pressure rise and volume flow rate.
(a) Outlet diameter is less than inlet diameter (Dout & Din) (b) Outlet and inlet diameters are equal (Dout ! Din) (c) Outlet diameter is greater than inlet diameter (Dout " Din) 14–9 An air compressor increases the pressure (Pout " Pin) and the density (rout " rin) of the air passing through it (Fig. P14–9). For the case in which the outlet and inlet diameters are equal (Dout ! Din), how does average air speed change across the compressor? In particular, is Vout less than, equal to, or greater than Vin? Explain. Answer: less than
14–3C List at least two common examples of fans, of blowers, and of compressors.
Din
14–4C Discuss the primary difference between a positivedisplacement turbomachine and a dynamic turbomachine. Give an example of each for both pumps and turbines.
rin, Vin Pin
14–5C For a pump, discuss the difference between brake horsepower and water horsepower, and also define pump efficiency in terms of these quantities. 14–6C For a turbine, discuss the difference between brake horsepower and water horsepower, and also define turbine efficiency in terms of these quantities. 14–7C Explain why there is an “extra” term in the Bernoulli equation in a rotating reference frame. 14–8 A water pump increases the pressure of the water passing through it (Fig. P14–8). The water is assumed to be incompressible. For each of the three cases listed below, how does average water speed change across the pump? In particular, is Vout less than, equal to, or greater than Vin? Show your equations, and explain. Dout Vout D in Vin
Pump
Pout
Pin
FIGURE P14–8 * Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them. Problems with the icon are solved using EES, and complete solutions together with parametric studies are included on the enclosed DVD. Problems with the icon are comprehensive in nature and are intended to be solved with a computer, preferably using the EES software that accompanies this text.
Dout Compressor
rout, Vout Pout
FIGURE P14–9 Pumps 14–10C There are three main categories of dynamic pumps. List and define them. 14–11C For each statement about centrifugal pumps, choose whether the statement is true or false, and discuss your answer briefly. (a) A centrifugal pump with radial blades has higher efficiency than the same pump with backward-inclined blades. (b) A centrifugal pump with radial blades produces a larger pressure rise than the same pump with. backward- or forwardinclined blades over a wide range of V. (c) A centrifugal pump with forward-inclined blades is a good choice when one needs to provide a large pressure rise over a wide range of volume flow rates. (d) A centrifugal pump with forward-inclined blades would most likely have less blades than a pump of the same size with backward-inclined or radial blades. 14–12C Figure P14–12C shows two possible locations for a water pump in a piping system that pumps water from the lower tank to the upper tank. Which location is better? Why? 14–13C Define net positive suction head and required net positive suction head, and explain how these two quantities are used to ensure that cavitation does not occur in a pump. 14–14C Consider flow through a water pump. For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) The faster the flow through the pump, the more likely that cavitation will occur.
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Reservoir
Reservoir
Valve
Pump
Valve Option (a)
Reservoir
Valve
14–17C Write the equation that defines actual (available) net positive suction head NPSH. From this definition, discuss at least five ways you can decrease the likelihood of cavitation in the pump, for the same liquid, temperature, and volume flow rate. 14–18C Consider steady, incompressible flow through two identical pumps (pumps 1 and 2), either in series or in parallel. For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) The volume . flow rate through the two pumps in series is . equal to V1 $ V2. (b) The overall net head across the two pumps in series is equal to H1 $ H2. (c) The volume flow rate through the two pumps in parallel . . is equal to V1 $ V2. (d) The overall net head across the two pumps in parallel is equal to H1 $ H2. 14–19C In Fig. P14–19C is shown a plot of the pump net head as a function of the pump volume flow rate, or capacity. On the figure, label the shutoff head, the free delivery, the pump performance curve, the system curve, and the operating point.
Reservoir Pump
H
Valve Option (b)
FIGURE P14–12C (b) As water temperature increases, NPSHrequired also increases. (c) As water temperature increases, the available NPSH also increases. (d) As water temperature increases, cavitation is less likely to occur. 14–15C Explain why it is usually not wise to arrange two (or more) dissimilar pumps in series or in parallel. 14–16C Consider a typical centrifugal liquid pump. For each statement, choose whether the statement is true or false, and discuss your answer briefly. . . (a) V at the pump’s free delivery is greater than V at its best efficiency point. (b) At the pump’s shutoff head, the pump efficiency is zero. (c) At the pump’s best efficiency point, its net head is at its maximum value. (d) At the pump’s free delivery, the pump efficiency is zero.
0
0
•
V
FIGURE P14–19C 14–20 Suppose the pump of Fig. P14–19C is situated between two water tanks with their free surfaces open to the atmosphere. Which free surface is at a higher elevation—the one corresponding to the tank supplying water to the pump inlet, or the one corresponding to the tank connected to the pump outlet? Justify your answer through use of the energy equation between the two free surfaces. 14–21 Suppose the pump of Fig. P14–19C is situated between two large water tanks with their free surfaces open to the atmosphere. Explain qualitatively what would happen to the pump performance curve if the free surface of the outlet tank were raised in elevation, all else being equal. Repeat for the system curve. What would happen to the operating point—would the volume flow rate at the operating point decrease, increase, or remain the same? Indicate the change . on a qualitative plot of H versus V, and discuss. (Hint: Use
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the energy equation between the free surface of the tank upstream of the pump and the free surface of the tank downstream of the pump.)
length of pipe is 6.7 m, and the elevation difference is (z1 # z2) ! 4.6 m. Estimate the volume flow rate through this piping system. Answer: 36.0 Lpm
14–22 Suppose the pump of Fig. P14–19C is situated between two large water tanks with their free surfaces open to the atmosphere. Explain qualitatively what would happen to the pump performance curve if a valve in the piping system were changed from 100 percent open to 50 percent open, all else being equal. Repeat for the system curve. What would happen to the operating point—would the volume flow rate at the operating point decrease, increase, or remain the . same? Indicate the change on a qualitative plot of H versus V, and discuss. (Hint: Use the energy equation between the free surface of the upstream tank and the free surface of the downstream tank.) Answer: decrease
14–25 Repeat Prob. 14–24, but with a rough pipe—pipe roughness e ! 0.50 mm. Assume that a modified pump is used, such that the new pump operates at its free delivery conditions, just as in Prob. 14–24. Assume all other dimensions and parameters are the same as in that problem. Do your results agree with intuition? Explain
14–23 Consider the flow system sketched in Fig. P14–23. The fluid is water, and the pump is a centrifugal pump. Generate a qualitative plot of the pump net head as a function of the pump capacity. On the figure, label the shutoff head, the free delivery, the pump performance curve, the system curve, and the operating point. (Hint: Carefully consider the required net head at conditions of zero flow rate.) 1
V1 ! 0
14–26
Consider the piping system of Fig. P14–23, with all the dimensions, parameters, minor loss coefficients, etc., of Prob. 14–24. The pump’s performance . follows a parabolic curve fit, Havailable ! H0 # aV 2, where H0 ! 17.6 m is the pump’s shutoff head, and a ! 0.00426 m/(Lpm)2 is a coefficient of. the curve fit. Estimate the operating volume flow rate V in Lpm (liters per minute), and compare with that of Prob. 14–24. Discuss. 14–27E The performance data for a centrifugal water pump are shown in Table P14–27E for water at 77°F (gpm ! gallons per minute). (a) For each row of data, calculate the pump efficiency (percent). Show all units and unit conversions for full credit. (b) Estimate the volume flow rate (gpm) and net head (ft) at the BEP of the pump.
z1
TA B L E P 1 4 – 2 7 E
Reservoir
Pump
2 z2
V2
FIGURE P14–23 14–24 Suppose the pump of Fig. P14–23 is operating at free delivery conditions. The pipe, both upstream and downstream of the pump, has an inner diameter of 2.0 cm and nearly zero roughness. The minor loss coefficient associated with the sharp inlet is 0.50, each valve has a minor loss coefficient of 2.4, and each of the three elbows has a minor loss coefficient of 0.90. The contraction at the exit reduces the diameter by a factor of 0.60 (60% of the pipe diameter), and the minor loss coefficient of the contraction is 0.15. Note that this minor loss coefficient is based on the average exit velocity, not the average velocity through the pipe itself. The total
. V, gpm
H, ft
bhp, hp
0.0 4.0 8.0 12.0 16.0 20.0 24.0
19.0 18.5 17.0 14.5 10.5 6.0 0.0
0.06 0.064 0.069 0.074 0.079 0.08 0.078
14–28 Transform each column of the. pump performance data of Prob. 14–27E to metric units: V into Lpm (liters per minute), H into m, and bhp into W. Calculate the pump efficiency (percent) using these metric values, and compare to those of Prob. 14–27E. 14–29E
For the centrifugal water pump of Prob. 14–27E, plot the pump’s performance. data: H (ft), bhp (hp), and hpump (percent) as functions of V (gpm), using symbols only (no lines). Perform linear least-squares polynomial curve fits for all three parameters, and plot the fitted curves as lines (no symbols) on the same plot. For con. sistency, use a first-order curve fit for H as a function of V 2., use a. second-order curve fit for bhp as a function of both V and .V 2,. and use. a third-order curve fit for hpump as a function of V, V 2, and V 3. List all curve-fitted equations and coefficients (with units) for full credit. Calculate the BEP of the pump based on the curve-fitted expressions.
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14–30E Suppose the pump of Probs. 14–27E and 14–29E is used in a piping system that . has the system requirement Hrequired ! (z2 # z1) $ bV 2, where elevation difference 2. z2 # z1 ! 15.5 ft, and coefficient b ! 0.00986 ft/(gpm) . Estimate the operating point of the system, namely, V operating (gpm) and Hoperating (ft). Answers: 9.14 gpm, 16.3 ft
cients, estimate the free delivery of the pump.] (b) The application requires 57.0 Lpm of flow at a pressure rise across the pump of 5.8 psi. Is this pump capable of meeting the requirements? Explain.
TA B L E P 1 4 – 3 4
14–31 The performance data for a centrifugal water pump are shown in Table P14–31 for water at 20°C (Lpm ! liters per minute). (a) For each row of data, calculate the pump efficiency (percent). Show all units and unit conversions for full credit. (b) Estimate the volume flow rate (Lpm) and net head (m) at the BEP of the pump.
TA B L E P 1 4 – 3 1 . V, Lpm
H, m
bhp, W
0.0 6.0 12.0 18.0 24.0 30.0 36.0
47.5 46.2 42.5 36.2 26.2 15.0 0.0
133 142 153 164 172 174 174
14–32
For the centrifugal water pump of Prob. 14–31, plot the pump’s performance . data: H (m), bhp (W), and hpump (percent) as functions of V (Lpm), using symbols only (no lines). Perform linear least-squares polynomial curve fits for all three parameters, and plot the fitted curves as lines (no symbols) on the same plot. For consistency, use a . first-order curve fit for H as a function of V 2, .use a second.2 order curve fit for bhp as a function of both V and V ., and . use a. third-order curve fit for hpump as a function of V , V 2, and V 3. List all curve-fitted equations and coefficients (with units) for full credit. Calculate the BEP of the pump based on the curve-fitted expressions. 14–33 Suppose the pump of Probs. 14–31 and 14–32 is used in a piping system .that has the system requirement Hrequired ! (z2 # z1) $ bV 2, where the elevation difference 2. z2 # z1 ! 10.0 m, and coefficient b ! 0.0185 m/(Lpm) . Estimate the operating point of the system, namely, V operating (Lpm) and Hoperating (m). 14–34
Suppose you are looking into purchasing a water pump with the performance data shown in Table P14–34. Your supervisor asks for some more information about the pump. (a) Estimate the shutoff head H0 and . the free delivery V max of the pump. [Hint: Perform a least. squares curve fit (regression analysis) of Havailable versus V 2, and calculate the best-fit values of coefficients H0 and a that translate the tabulated data of Table P14–34 into the para. bolic expression, Havailable ! H0 # aV 2. From these coeffi-
. V, Lpm
H, m
20 30 40 50
21 18.4 14 7.6
14–35E A manufacturer of small water pumps lists the performance data for a family . of its pumps as a parabolic curve fit, Havailable ! H0 # aV 2, where H0 is the pump’s shutoff head and a is a coefficient. Both H0 and a are listed in a table for the pump family, along with the pump’s free delivery. The pump head is given in units of feet of water column, and capacity is given in units of gallons per minute. (a) What are the units of coefficient. a? (b) Generate an expression for the pump’s free delivery V max in terms of H0 and a. (c) Suppose one of the manufacturer’s pumps is used to pump water from one large reservoir to another at a higher elevation. The free surfaces of both reservoirs are exposed to atmospheric pressure. . The system curve simplifies to Hrequired ! (z. 2 # z1) $ bV 2. Calculate the operating point of the pump (V operating and Hoperating) in terms of H0, a, b, and elevation difference z2 # z1. 14–36 The performance data. of a water pump follow the curve fit Havailable ! H0 # aV 2, where the pump’s shutoff 2, the head H0 ! 5.30 m, coefficient a ! 0.0453 m/(Lpm) . units of pump head H are meters, and the units of V are liters per minute (Lpm). The pump is used to pump water from one large reservoir to another large reservoir at a higher elevation. The free surfaces of both reservoirs are exposed to atmospheric pressure. The system curve simplifies to Hrequired ! (z2 . # z1) $ bV 2, where elevation difference z2 # z1 ! 3.52 m, and coefficient b ! 0.0261 m(Lpm)2. Calculate the operating . point of the pump (V operating and Hoperating) in appropriate units (Lpm and meters, respectively). Answers: 4.99 Lpm, 4.17 m 14–37E A water pump is used to pump water from one large reservoir to another large reservoir that is at a higher elevation. The free surfaces of both reservoirs are exposed to atmospheric pressure, as sketched in Fig. P14–37E. The dimensions and minor loss coefficients are provided in the figure. The pump’s performance . is approximated by the expression Havailable ! H0 # aV 2, where the shutoff head H0 ! 125 ft of water column, coefficient a ! 2.50 ft/gpm2, available pump head. Havailable is in units of feet of water column, and capacity V is in units of gallons per minute (gpm). Estimate the capacity delivered by the pump. Answer: 6.34 gpm
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808 FLUID MECHANICS z2 – z1 = 22.0 ft (elevation difference) D = 1.20 in (pipe diameter) KL, entrance = 0.50 (pipe entrance) KL, valve 1 = 2.0 (valve 1) KL, valve 2 = 6.8 (valve 2) KL, elbow = 0.34 (each elbow—there are 3) KL, exit = 1.05 (pipe exit) L = 124 ft (total pipe length) e = 0.0011 in (pipe roughness) 2 z2
V2 ! 0
Reservoir
replaced with the new pump. (a) Perform a least-squares . curve fit (regression analysis) of Havailable versus V 2, and calculate the best-fit values of coefficients H0 and a that translate the tabulated data of Table . P14–40E into the parabolic expression Havailable ! H0 # aV 2. Plot the data points as symbols and the curve fit as a line for comparison. (b) Estimate the operating volume flow rate of the new pump if it were to replace the existing pump, all else being equal. Compare to the result of Prob. 14–37E and discuss. Is Paul correct? (c) Generate a plot of required net head and available net head as functions of volume flow rate and indicate the operating point on the plot.
z2 – z1
TA B L E P 1 4 – 4 0 E D
1
V1 ! 0
z1
Reservoir Valve 2
Pump Valve 1
FIGURE P14–37E 14–38E For the pump and piping system of Prob. 14–37E, plot the required pump head H. required (ft of water column) as a function of volume flow rate V (gpm). On the same plot, com. pare the available pump head Havailable versus V , and mark the operating point. Discuss. 14–39E Suppose that the two reservoirs in Prob. 14–37E are 1000 ft further apart horizontally, but at the same elevations. All the constants and parameters are identical to those of Prob. 14–37E except that the total pipe length is 1124 ft instead of 124 ft. Calculate the volume flow rate for this case and compare with the result of Prob. 14–37E. Discuss. 14–40E
Paul realizes that the pump being used in Prob. 14–37E is not well-matched for this application, since its shutoff head (125 ft) is much larger than its required net head (less than 30 ft), and its capacity is fairly low. In other words, this pump is designed for highhead, low-capacity applications, whereas the application at hand is fairly low-head, and a higher capacity is desired. Paul tries to convince his supervisor that a less expensive pump, with lower shutoff head but higher free delivery, would result in a significantly increased flow rate between the two reservoirs. Paul looks through some online brochures, and finds a pump with the performance data shown in Table P14–40E. His supervisor asks him to predict the volume flow rate between the two reservoirs if the existing pump were
. V, gpm
H, ft
0 4 8 12 16 20 24
38 37 34 29 21 12 0
14–41 A water pump is used to pump water from one large reservoir to another large reservoir that is at a higher elevation. The free surfaces of both reservoirs are exposed to atmospheric pressure, as sketched in Fig. P14–41. The dimensions and minor loss coefficients are provided in the z2 – z1 = 7.85 m (elevation difference) D = 2.03 cm (pipe diameter) KL, entrance = 0.50 (pipe entrance) KL, valve = 17.5 (valve) KL, elbow = 0.92 (each elbow—there are 5) KL, exit = 1.05 (pipe exit) L = 176.5 m (total pipe length) e = 0.25 mm (pipe roughness) 2 z2
V2 ! 0
Reservoir z2 – z1
1
V1 ! 0
D
z1
Reservoir Pump Valve
FIGURE P14–41
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809 CHAPTER 14
figure. The pump’s performance . is approximated by the expression Havailable ! H0 # aV 2, where shutoff head H0 ! 24.4 m of water column, coefficient a ! 0.0678 m/Lpm2, available pump head H. available is in units of meters of water column, and capacity V is in units of liters per minute (Lpm). Estimate the capacity delivered by the pump. Answer: 11.6 Lpm
14–42 For the pump and piping system of Prob. 14–41, plot (m of water column) as a function required pump head Hrequired . of volume flow rate V (Lpm). On the . same plot, compare available pump head Havailable versus V , and mark the operating point. Discuss. 14–43 Suppose that the free surface of the inlet reservoir in Prob. 14–41 is 5.0 m higher in elevation, such that z2 # z1 ! 2.85 m. All the constants and parameters are identical to those of Prob. 14–41 except for the elevation difference. Calculate the volume flow rate for this case and compare with the result of Prob. 14–41. Discuss. 14–44
April’s supervisor asks her to find a replacement pump that will increase the flow rate through the piping system of Prob. 14–41 by a factor of 2 or greater. April looks through some online brochures, and finds a pump with the performance data shown in Table P14–44. All dimensions and parameters remain the same as in Prob. 14–41—only the pump is changed. (a) Perform a least. squares curve fit (regression analysis) of Havailable versus V 2, and calculate the best-fit values of coefficients H0 and a that translate the tabulated data of Table . P14–44 into the parabolic expression Havailable ! H0 # aV 2. Plot the data points as symbols and the curve fit as a line for comparison. (b) Use the expression obtained in part (a) to estimate the operating volume flow rate of the new pump if it were to replace the existing pump, all else being equal. Compare to the result of Prob. 14–41 and discuss. Has April achieved her goal? (c) Generate a plot of required net head and available net head as functions of volume flow rate, and indicate the operating point on the plot.
14–46 Comparing the results of Probs. 14–41 and 14–45, the volume flow rate increases as expected when one doubles the inner diameter of the pipe. One might expect that the Reynolds number increases as well. Does it? Explain. 14–47 Repeat Prob. 14–41, but neglect all minor losses. Compare the volume flow rate with that of Prob. 14–41. Are minor losses important in this problem? Discuss. 14–48
Consider the pump and piping system of Prob. 14–41. Suppose that the lower reservoir is huge, and its surface does not change elevation, but the upper reservoir is not so big, and its surface rises slowly. as the reservoir fills. Generate a curve of volume flow rate V (Lpm) as a function of z2 # z1 in the range 0 to the value of z2 # z1 at which the pump ceases to pump any more water. At what value of z2 # z1 does this occur? Is the curve linear? Why or why not? What would happen if z2 # z1 were greater than this value? Explain. 14–49E A local ventilation system (a hood and duct system) is used to remove air and contaminants produced by a welding operation (Fig. P14–49E). The inner diameter (ID) of the duct is D ! 9.06 in, its average roughness is 0.0059 in, and its total length is L ! 34.0 ft. There are three elbows along the duct, each with a minor loss coefficient of 0.21. Literature from the hood manufacturer lists the hood entry loss coefficient as 4.6 based on duct velocity. When the damper is fully open, its loss coefficient is 1.8. A squirrel
z2 2
TA B L E P 1 4 – 4 4 . V, Lpm
H, m
0 5 10 15 20 25 30
46.5 46 42 37 29 16.5 0
14–45 Calculate the volume flow rate between the reservoirs of Prob. 14–41 for the case in which the pipe diameter is doubled, all else remaining the same. Discuss.
Fan
Damper
Hood
1
FIGURE P14–49E
z1
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810 FLUID MECHANICS
cage centrifugal fan with a 9.0-in inlet is available. Its performance data fit a parabolic curve of the form Havailable ! H0 # . aV 2, where shutoff head H0 ! 2.30 inches of water column, coefficient a ! 8.50 % 10#6 inches of water column per (SCFM)2, available head H.available is in units of inches of water column, and capacity V is in units of standard cubic feet per minute (SCFM, at 77°F). Estimate the volume flow rate in SCFM through this ventilation system. Answer: 452 SCFM 14–50E For the duct system and fan of Prob. 14–49E, partially closing the damper would decrease the flow rate. All else being unchanged, estimate the minor loss coefficient of the damper required to decrease the volume flow rate by a factor of 2. 14–51E Repeat Prob. 14–49E, ignoring all minor losses. How important are the minor losses in this problem? Discuss. 14–52 A local ventilation system (a hood and duct system) is used to remove air and contaminants from a pharmaceutical lab (Fig. P14–52). The inner diameter (ID) of the duct is D ! 150 mm, its average roughness is 0.15 mm, and its total length is L ! 24.5 m. There are three elbows along the duct, each with a minor loss coefficient of 0.21. Literature from the hood manufacturer lists the hood entry loss coefficient as 3.3 based on duct velocity. When the damper is fully open, its loss coefficient is 1.8. The minor loss coefficient through the 90° tee is 0.36. Finally, a one-way valve is installed to prevent contaminants from a second hood from flowing “backward” into the room. The minor loss coefficient of the (open)
2
z2
90° Tee
One-way valve
one-way valve is 6.6. The performance data of the. fan fit a parabolic curve of the form Havailable ! H0 # aV 2, where shutoff head H0 ! 60.0 mm of water column, coefficient a ! 2.50 % 10#7 mm of water column per (Lpm)2, available head Havailable is in units of mm of water column, and capacity . V is in units of Lpm of air. Estimate the volume flow rate in Lpm through this ventilation system. Answer: 7090 Lpm 14–53 For the duct system of Prob. 14–52, plot required fan head Hrequired . (mm of water column) as a function of volume flow rate V (Lpm). On . the same plot, compare available fan head Havailable versus V , and mark the operating point. Discuss. 14–54 Repeat Prob. 14–52, ignoring all minor losses. How important are the minor losses in this problem? Discuss. 14–55 Suppose the one-way valve of Fig. P14–52 malfunctions due to corrosion and is stuck in its fully closed position (no air can get through). The fan is on, and all other conditions are identical to those of Prob. 14–52. Calculate the gage pressure (in pascals and in mm of water column) at a point just downstream of the fan. Repeat for a point just upstream of the one-way valve. 14–56E A centrifugal pump is used to pump water at 77°F from a reservoir whose surface is 20.0 ft above the centerline of the pump inlet (Fig. P14–56E). The piping system consists of 67.5 ft of PVC pipe with an ID of 1.2 in and negligible average inner roughness height. The length of pipe from the bottom of the lower reservoir to the pump inlet is 12.0 ft. There are several minor losses in the piping system: a sharpedged inlet (KL ! 0.5), two flanged smooth 90° regular elbows (KL ! 0.3 each), two fully open flanged globe valves (KL ! 6.0 each), and an exit loss into the upper reservoir (KL ! 1.05). The pump’s required net positive suction head is provided by the manufacturer . as a curve fit: NPSHrequired ! 1.0 ft $ (0.0054 ft/gpm2)V 2, where volume flow rate is in gpm. Estimate the maximum volume flow rate (in units of gpm) that can be pumped without cavitation.
Branch from another hood
3
z3 Reservoir
Fan z3 – z1 1
Damper
z1
Reservoir
Valve 2
Hood
z2 1
FIGURE P14–52
z1
Valve
FIGURE P14–56E
Pump
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811 CHAPTER 14
14–57E Repeat Prob. 14–56E, but at a water temperature of 150°F. Discuss. 14–58 A self-priming centrifugal pump is used to pump water at 25°C from a reservoir whose surface is 2.2 m above the centerline of the pump inlet (Fig. P14–58). The pipe is PVC pipe with an ID of 24.0 mm and negligible average inner roughness height. The pipe length from the submerged pipe inlet to the pump inlet is 2.8 m. There are only two minor losses in the piping system from the pipe inlet to the pump inlet: a sharp-edged reentrant inlet (KL ! 0.85), and a flanged smooth 90° regular elbow (KL ! 0.3). The pump’s required net positive suction head is provided by the manufacturer as. a curve fit: NPSHrequired ! 2.2 m $ (0.0013 m/Lpm2)V 2, where volume flow rate is in Lpm. Estimate the maximum volume flow rate (in units of Lpm) that can be pumped without cavitation.
Pump z2 2 1
z2 – z1 z1
Reservoir
FIGURE P14–58 14–59 Repeat Prob. 14–58, but at a water temperature of 80°C. Repeat for 90°C. Discuss. 14–60 Repeat Prob. 14–58, but with the pipe diameter increased by a factor of 2 (all else being equal). Does the volume flow rate at which cavitation occurs in the pump increase or decrease with the larger pipe? Discuss. 14–61 Two water pumps are arranged in series. The performance data for both . pumps follow the parabolic curve fit Havailable ! H0 # aV 2. For pump 1, H0 ! 5.30 m and coefficient a ! 0.0438 m/Lpm2; for pump 2, H0 ! 7.80 m and coefficient a ! 0.0347 m/Lpm2. In either case, the . units of net pump head H are m, and the units of capacity V are Lpm. Calculate the combined shutoff head and free delivery of the two pumps working together in series. At what volume flow rate should pump 1 be shut off and bypassed? Explain. Answers: 13.1 m, 15.0 Lpm, 11.0 Lpm
14–62 The same two water pumps of Prob. 14–61 are arranged in parallel. Calculate the shutoff head and free delivery of the two pumps working together in parallel. At what combined net head should pump 1 be shut off and bypassed? Explain.
14–63E The two-lobe rotary pump of Fig. P14–63E moves . 0.145 gal of a coal slurry in each lobe volume V lobe. Calculate the volume flow rate of the slurry (in gpm) for the case . where n ! 300 rpm. Answer: 174 gpm
⋅ V
In
⋅ V
Out
Vlobe
FIGURE P14–63E 14–64E Repeat Prob. 14–63E for the case in which .the pump has three lobes on each rotor instead of two, and V lobe ! 0.087 gal. 14–65 A two-lobe rotary positive-displacement pump, similar to that of Fig. 14–30, moves 3.64 cm3 of tomato paste in . each lobe volume V lobe. Calculate the volume flow rate of . tomato paste for the case where n ! 400 rpm. 14–66 Consider the gear pump of Fig. 14–26c. Suppose the volume of fluid confined between two gear teeth is 0.350 cm3. How much fluid volume is pumped per rotation? Answer: 9.80 cm3
. 14–67 A centrifugal pump rotates at n ! 750 rpm. Water enters the impeller normal to the blades (a1 ! 0°) and exits at an angle of 35° from radial (a2 ! 35°). The inlet radius is r1 ! 12.0 cm, at which the blade width b1 ! 18.0 cm. The outlet radius is r2 ! 24.0 cm, at which the blade width b2 ! 14.0 cm. The volume flow rate is 0.573 m3/s. Assuming 100 percent efficiency, calculate the net head produced by this pump in cm of water column height. Also calculate the required brake horsepower in W. 14–68 A vane-axial flow fan is being designed with the stator blades upstream of the rotor blades (Fig. P14–68). To Hub and motor vr
bst ??? r
→
Vin Stator
FIGURE P14–68
v
Rotor
→
Vout
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812 FLUID MECHANICS
reduce expenses, both the stator and rotor blades are to be constructed of sheet metal. The stator blade is a simple circular arc with its leading edge aligned axially and its trailing edge at angle bst ! 26.6° from the axis as shown in the sketch. (The subscript notation indicates stator trailing edge.) There are 18 stator blades. At design conditions, the axialflow speed through the blades is 31.4 m/s, and the impeller rotates at 1800 rpm. At a radius of 0.50 m, calculate the leading and trailing edge angles of the rotor blade, and sketch the shape of the blade. How many rotor blades should there be?
Turbines 14–69C Give at least two reasons why turbines often have greater efficiencies than do pumps.
itself has an efficiency of 94.5 percent. Estimate the electric power production from the plant in MW. 14–78 A Francis radial-flow hydroturbine is being designed with the following dimensions: r2 ! 2.00 m, r1 ! 1.42 m, b2 . ! 0.731 m, and b1 ! 2.20 m. The runner rotates at n ! 180 rpm. The wicket gates turn the flow by angle a2 ! 30° from radial at the runner inlet, and the flow at the runner outlet is at angle a1 ! 10° from radial (Fig. P14–78). The volume flow rate at design conditions is 340 m3/s, and the gross head provided by the dam is Hgross ! 90.0 m. For the preliminary design, irreversible losses are neglected. Calculate the inlet and outlet runner blade angles b2 and b1, respectively, and predict the power output (MW) and required net head (m). Is the design feasible?
14–70C Briefly discuss the main difference in the way that dynamic pumps and reaction turbines are classified as centrifugal (radial), mixed flow, or axial.
V2, t a2
14–71C What is a draft tube, and what is its purpose? Describe what would happen if turbomachinery designers did not pay attention to the design of the draft tube.
→
V2
14–72C Name and briefly describe the differences between the two basic types of dynamic turbine.
14–76 Some engineers are evaluating potential sites for a small hydroelectric dam. At one such site, the gross head is 650 m, and they estimate that the volume flow rate of water through each turbine would be 1.5 m3/s. Estimate the ideal power production per turbine in MW. 14–77E A hydroelectric power plant is being designed. The gross head from the reservoir to the tailrace is 1065 ft, and the volume flow rate of water through each turbine is 203,000 gpm at 70°F. There are 12 identical parallel turbines, each with an efficiency of 95.2 percent, and all other mechanical energy losses (through the penstock, etc.) are estimated to reduce the output by 3.5 percent. The generator
→
a1
V1
V1, n
r1 v
14–74 Prove that for a given jet speed, volume flow rate, turning angle, and wheel radius, the maximum shaft power produced by a Pelton wheel occurs when the turbine bucket moves at half the jet speed.
Answers: (a) 0.801 m3/s, (b) 266 rpm, (c) 3.35 MW
V1, t
r2
14–73C Discuss the meaning of reverse swirl in reaction hydroturbines, and explain why some reverse swirl is desirable. Use an equation to support your answer. Why is it not wise to have too much reverse swirl?
14–75 A Pelton wheel is used to produce hydroelectric power. The average radius of the wheel is 1.83 m, and the jet velocity is 102 m/s from a nozzle of exit diameter equal to 10.0 cm. The turning angle of the buckets is b ! 165°. (a) Calculate the volume flow rate through the turbine in m3/s. (b) What is the optimum rotation rate (in rpm) of the wheel (for maximum power)? (c) Calculate the output shaft power in MW if the efficiency of the turbine is 82 percent.
V2, n
Control volume
FIGURE P14–78 14–79
Reconsider Prob. 14–78. Using EES (or other) software, investigate the effect of the runner outlet angle a1 on the required net head and the output power. Let the outlet angle vary from #20° to 20° in increments of 1°, and plot your results. Determine the minimum possible value of a1 such that the flow does not violate the laws of thermodynamics. 14–80 A Francis radial-flow hydroturbine has the following dimensions, where location 2 is the inlet and location 1 is the outlet: r2 ! 2.00 m, r1 ! 1.30 m, b2 ! 0.85 m, and b1 ! 2.10 m. The runner blade angles are b2 ! 66° and b1 ! 18.5° at the turbine inlet and outlet, respectively. The run. ner rotates at n ! 100 rpm. The volume flow rate at design conditions is 80.0 m3/s. Irreversible losses are neglected in this preliminary analysis. Calculate the angle a2 through which the wicket gates should turn the flow, where a2 is measured from the radial direction at the runner inlet (Fig.
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813 CHAPTER 14
P14–78). Calculate the swirl angle a1, where a1 is measured from the radial direction at the runner outlet (Fig. P14–78). Does this turbine have forward or reverse swirl? Predict the power output (MW) and required net head (m).
14–88 Consider the fan of Prob. 14–52. The fan diameter is . 30.0 cm, and it operates at n ! 600 rpm. Nondimensionalize the pump performance curve, i.e., plot . CH versus CQ. Show sample calculations of CH and CQ at V ! 13,600 Lpm.
14–81E A Francis radial-flow hydroturbine has the following dimensions, where location 2 is the inlet and location 1 is the outlet: r2 ! 6.60 ft, r1 ! 4.40 ft, b2 ! 2.60 ft, and b1 ! 7.20 ft. The runner blade angles are b2 ! 82° and b1 ! 46° at the turbine inlet and outlet, respectively. The runner rotates . at n ! 120 rpm. The volume flow rate at design conditions is 4.70 % 106 gpm. Irreversible losses are neglected in this preliminary analysis. Calculate the angle a2 through which the wicket gates should turn the flow, where a2 is measured from the radial direction at the runner inlet (Fig. P14–78). Calculate the swirl angle a1, where a1 is measured from the radial direction at the runner outlet (Fig. P14–78). Does this turbine have forward or reverse swirl? Predict the power output (hp) and required net head (ft).
14–89 Calculate the pump specific speed of the fan of Prob. 14–88 at the best efficiency point for the case in which the BEP occurs at 13,600 Lpm. Provide answers in both dimensionless form and in customary U.S. units. What kind of fan is it?
14–82E
Using EES or other software, adjust the runner blade trailing edge angle b1 of Prob. 14–81E, keeping all other parameters the same, such that there is no swirl at the turbine outlet. Report b1 and the corresponding shaft power.
Pump and Turbine Scaling Laws 14–83C Look up the word affinity in a dictionary. Why do you suppose some engineers refer to the turbomachinery scaling laws as affinity laws? 14–84C For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) If the rpm of a pump is doubled, all else staying the same, the capacity of the pump goes up by a factor of about 2. (b) If the rpm of a pump is doubled, all else staying the same, the net head of the pump goes up by a factor of about 2. (c) If the rpm of a pump is doubled, all else staying the same, the required shaft power goes up by a factor of about 4. (d) If the rpm of a turbine is doubled, all else staying the same, the output shaft power of the turbine goes up by a factor of about 8. 14–85C Discuss which dimensionless pump performance parameter is typically used as the independent parameter. Repeat for turbines instead of pumps. Explain. 14–86 Consider the pump of Prob. 14–41. The pump diam. eter is 1.80 cm, and it operates at n ! 4200 rpm. Nondimensionalize the pump performance curve, i.e., plot . CH versus CQ. Show sample calculations of CH and CQ at V ! 14.0 Lpm. 14–87 Calculate the pump specific speed of the pump of Prob. 14–86 at the best efficiency point for the case in which the BEP occurs at 14.0 Lpm. Provide answers in both dimensionless form and in customary U.S. units. What kind of pump is it? Answers: 0.199, 545, centrifugal
14–90 Calculate the pump specific speed of the pump of Example 14–11 at its best efficiency point. Provide answers in both dimensionless form and in customary U.S. units. What kind of pump is it? 14–91 Len is asked to design a small water pump for an aquarium. The pump should deliver 18.0 Lpm of water at a net head of 1.6 m at its best efficiency point. A motor that spins at 1200 rpm is available. What kind of pump (centrifugal, mixed, or axial) should Len design? Show all your calculations and justify your choice. Estimate the maximum pump efficiency Len can hope for with this pump. Answers: centrifugal, 75%
14–92E A large water pump is being designed for a nuclear reactor. The pump should deliver 2500 gpm of water at a net head of 45 ft at its best efficiency point. A motor that spins at 300 rpm is available. What kind of pump (centrifugal, mixed, or axial) should be designed? Show all your calculations and justify your choice. Estimate the maximum pump efficiency that can be hoped for with this pump. Estimate the power (brake horsepower) required to run the pump. 14–93 Consider the pump of Prob. 14–91. Suppose the pump is modified by attaching a different motor, for which the rpm is half that of the original pump. If the pumps operate at homologous points (namely, at the BEP) for both cases, predict the volume flow rate and net head of the modified pump. Calculate the pump specific speed of the modified pump, and compare to that of the original pump. Discuss. 14–94 Verify that turbine specific speed and pump specific speed are related as follows: N St ! N Sp 1h turbine. 14–95 Consider a pump–turbine that operates both as a pump and as a turbine. Under conditions in . which the rotational speed v and the volume flow rate V are the same for the pump and the turbine, verify that turbine specific speed and pump specific speed are related as Hpump 3*4 N St ! N Sp 2h turbine a b Hturbine
bhppump 3*4 ! N Sp(h turbine)5*4(h pump)3*4 a b bhpturbine
14–96 Apply the necessary conversion factors to prove the relationship between dimensionless turbine specific speed and conventional U.S. turbine specific speed, NSt ! 43.46NSt, US.
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814 FLUID MECHANICS
Note that we assume water as the fluid and standard earth gravity. 14–97 Calculate the turbine specific speed of the Round Butte hydroturbine of Fig 14–89. Does it fall within the range of NSt appropriate for that type of turbine? 14–98 Calculate the turbine specific speed of the Smith Mountain hydroturbine of Fig 14–90. Does it fall within the range of NSt appropriate for that type of turbine? 14–99 Calculate the turbine specific speed of the Warwick hydroturbine of Fig 14–91. Does it fall within the range of NSt appropriate for that type of turbine? 14–100 Calculate the turbine specific speed of the turbine of Example 14–12 for the case where a1 ! 10°. Provide answers in both dimensionless form and in customary U.S. units. Is it in the normal range for a Francis turbine? If not, what type of turbine would be more appropriate? 14–101 Calculate the turbine specific speed of the turbine in Prob. 14–80. Provide answers in both dimensionless form and in customary U.S. units. Is it in the normal range for a Francis turbine? If not, what type of turbine would be more appropriate? 14–102E Calculate the turbine specific speed of the turbine in Prob. 14–81E using customary U.S. units. Is it in the normal range for a Francis turbine? If not, what type of turbine would be more appropriate? 14–103 Calculate the turbine specific speed of the turbine in Prob. 14–78. Provide answers in both dimensionless form and in customary U.S. units. Is it in the normal range for a Francis turbine? If not, what type of turbine would be more appropriate? 14–104 A one-fifth scale model of a water turbine is tested in a laboratory at T ! 20°C. The diameter of the model is 8.0 cm, its volume flow rate is 17.0 m3/h, it spins at 1500 rpm, and it operates with a net head of 15.0 m. At its best efficiency point, it delivers 450 W of shaft power. Calculate the efficiency of the model turbine. What is the most likely kind of turbine being tested? Answers: 64.9%, impulse 14–105 The prototype turbine corresponding to the onefifth scale model turbine discussed in Prob. 14–104 is to operate across a net head of 50 m. Determine the appropriate rpm and volume flow rate for best efficiency. Predict the brake horsepower output of the prototype turbine, assuming exact geometric similarity. 14–106 Prove that the model turbine (Prob. 14–104) and the prototype turbine (Prob. 14–105) operate at homologous points by comparing turbine efficiency and turbine specific speed for both cases. 14–107 In Prob. 14–106, we scaled up the model turbine test results to the full-scale prototype assuming exact dynamic similarity. However, as discussed in the text, a large
prototype typically yields higher efficiency than does the model. Estimate the actual efficiency of the prototype turbine. Briefly explain the higher efficiency. 14–108 A group of engineers is designing a new hydroturbine by scaling up an existing one. The existing turbine . (turbine A) has diameter DA ! 1.50 m,. and spins at nA 3 ! 150 rpm. At its best efficiency point, V A ! 162 m /s, HA ! 90.0 m of water, and bhpA ! 132 MW. The new turbine (turbine B) will spin at 120 rpm, and its net head will be HB ! 110 m. Calculate the diameter of the new turbine such that . it operates most efficiently, and calculate V B and bhpB. Answers: 2.07 m, 342 m3/s, 341 MW
14–109 Calculate and compare the efficiency of the two turbines of Prob. 14–108. They should be the same since we are assuming dynamic similarity. However, the larger turbine will actually be slightly more efficient than the smaller turbine. Use the Moody efficiency correction equation to predict the actual expected efficiency of the new turbine. Discuss. 14–110 Calculate and compare the turbine specific speed for both the small (A) and large (B) turbines of Prob. 14–108. What kind of turbine are these most likely to be?
Review Problems 14–111C For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) A gear pump is a type of positive-displacement pump. (b) A rotary pump is a type of positive-displacement pump. (c) The pump performance curve (net head versus capacity) of a positive-displacement pump is nearly vertical throughout its recommended operating range at a given rotational speed. (d) At a given rotational speed, the net head of a positivedisplacement pump decreases with fluid viscosity. 14–112C The common water meter found in most homes can be thought of as a type of turbine, since it extracts energy from the flowing water to rotate the shaft connected to the volume-counting mechanism (Fig. P14–112C). From the point of view of a piping system, however (Chap. 8), what kind of device is a water meter? Explain.
Water meter
FIGURE P14–112C 14–113C Pump specific speed and turbine specific speed are “extra” parameters that are not necessary in the scaling laws for pumps and turbines. Explain, then, their purpose. 14–114C What is a pump–turbine? Discuss an application where a pump–turbine is useful.
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815 CHAPTER 14
14–115 For two dynamically similar pumps, manipulate the dimensionless parameters to show that DB . pump . ! DA(HA/HB)1/4(V B/V A)1/2. Does the same relationship apply to two dynamically similar turbines? 14–116 For two dynamically similar turbines, manipulate the dimensionless turbine parameters to show that DB ! DA(HA/HB)3/4(rA/rB)1/2(bhpB/bhpA)1/2. Does the same relationship apply to two dynamically similar pumps?
Design and Essay Problems 14–117
Develop a general-purpose computer application (using EES or other software) that employs the affinity laws to design a new pump (B) that is dynamically similar to a given pump (A). The inputs for pump A are diameter, net head, capacity, density, rotational speed, and pump efficiency. The inputs for pump B are density (rB may differ from rA), desired net head, and desired capacity. The outputs for pump B are diameter, rotational speed, and required shaft power. Test your program using . the following inputs: DA ! 5.0 cm, HA ! 120 cm, V A . ! 400 cm3/s, rA ! 998.0 kg/m3, nA ! 1725 rpm, hpump,.A 3 ! 81 percent, rB ! 1226 kg/m , HB ! 450 cm, and V B ! 2400 cm3/s. Verify your results manually. Answers: DB . ! 8.80 cm, n B ! 1898 rpm, and bhpB ! 160 W
14–118
Experiments on an existing pump (A) yield .the following BEP data: DA ! 10.0 cm, H.A ! 210 cm, V A ! 1350 cm3/s, rA ! 998.0 kg/m3, nA ! 1500 rpm, hpump, A ! 87 percent. You are to design a new pump (B) that has the following requirements: rB . ! 998.0 kg/m3, HB ! 570 cm, and V B ! 3670 cm3/s. Apply the computer program you developed in Prob. 14–117 to cal. culate DB (cm), nB (rpm), and bhpB (W). Also calculate the pump specific speed. What kind of pump is this (most likely)?
14–119
Develop a general-purpose computer application (using EES or other software) that employs the affinity laws to design a new turbine (B) that is dynamically similar to a given turbine (A). The inputs for turbine A are diameter, net head, capacity, density, rotational speed, and brake horsepower. The inputs for turbine B are density (rB may differ from rA), available net head, and rotational speed. The outputs for turbine B are diameter, capacity, and brake horsepower. Test your program .using the following inputs: DA ! 1.40 m, HA ! 80.0 m, V A ! 162 m3/s, rA . ! 998.0 kg/m3, nA ! 150 rpm, bhpA ! 118 MW, rB . ! 998.0 kg/m3, HB ! 95.0 m, and nB ! 120 rpm. Verify your . results manually. Answers: DB ! 1.91 m, V B ! 328 m3/s, and bhpB ! 283 MW
14–120
Experiments on an existing turbine (A) yield the following data: DA ! 86.0 cm, HA . . ! 22.0 m, V A ! 69.5 m3/s, rA ! 998.0 kg/m3, nA ! 240 rpm, bhpA ! 11.4 MW. You are to design a new turbine (B) that has the following requirements: rB ! 998.0 kg/m3, HB . ! 95.0 m, and nB ! 210 rpm. Apply the computer program . you developed in Prob. 14–119 to calculate DB (m), V B (m3/s), and bhpB (MW). Also calculate the turbine specific speed. What kind of turbine is this (most likely)? 14–121 Calculate and compare the efficiency of the two turbines of Prob. 14–120. They should be the same since we are assuming dynamic similarity. However, the larger turbine will actually be slightly more efficient than the smaller turbine. Use the Moody efficiency correction equation to predict the actual expected efficiency of the new turbine. Discuss.
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CHAPTER
15
INTRODUCTION T O C O M P U T AT I O N A L FLUID DYNAMICS
A
brief introduction to computational fluid dynamics (CFD) is presented in this chapter. While any intelligent, computer-literate person can run a CFD code, the results he or she obtains may not be physically correct. In fact, if the grid is not properly generated, or if the boundary conditions or flow parameters are improperly applied, the results may even be completely erroneous. Therefore, the goal of this chapter is to present guidelines about how to generate a grid, how to specify boundary conditions, and how to determine if the computer output is meaningful. We stress the application of CFD to engineering problems, rather than details about grid generation techniques, discretization schemes, CFD algorithms, or numerical stability. The examples presented here have been obtained with the commercial computational fluid dynamics code FLUENT. Other CFD codes would yield similar, but not identical results. Sample CFD solutions are shown for incompressible and compressible laminar and turbulent flows, flows with heat transfer, and flows with free surfaces. As always, one learns best by hands-on practice. For this reason, we provide several homework problems that utilize FLUENT FLOWLAB®, hereafter referred to as FlowLab, a student-friendly CFD code that is available with the purchase of this book.
OBJECTIVES When you finish reading this chapter, you should be able to ■ Understand the importance of a high-quality, good resolution mesh ■
■
■
Apply appropriate boundary conditions to computational domains Understand how to apply CFD to basic engineering problems and how to determine whether the output is physically meaningful Realize that you need much further study and practice to use CFD successfully
817
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818 FLUID MECHANICS
15–1
■
INTRODUCTION AND FUNDAMENTALS
Motivation
FIGURE 15–1 CFD calculations of the ascent of the space shuttle launch vehicle (SSLV). The grid consists of more than 16 million points, and filled pressure contours are shown. Free-stream conditions are Ma " 1.25, and the angle of attack is #3.3°. NASA/Photo by Ray J. Gomez. Used by permission.
There are two fundamental approaches to design and analysis of engineering systems that involve fluid flow: experimentation and calculation. The former typically involves construction of models that are tested in wind tunnels or other facilities (Chap. 7), while the latter involves solution of differential equations, either analytically (Chaps. 9 and 10) or computationally. In the present chapter, we provide a brief introduction to computational fluid dynamics (CFD), the field of study devoted to solution of the equations of fluid flow through use of a computer (or, more recently, several computers working in parallel). Modern engineers apply both experimental and CFD analyses, and the two complement each other. For example, engineers may obtain global properties, such as lift, drag, pressure drop, or power, experimentally, but use CFD to obtain details about the flow field, such as shear stresses, velocity and pressure profiles (Fig. 15–1), and flow streamlines. In addition, experimental data are often used to validate CFD solutions by matching the computationally and experimentally determined global quantities. CFD is then employed to shorten the design cycle through carefully controlled parametric studies, thereby reducing the required amount of experimental testing. The current state of computational fluid dynamics is that CFD can handle laminar flows with ease, but turbulent flows of practical engineering interest are impossible to solve without invoking turbulence models. Unfortunately, no turbulence model is universal, and a turbulent CFD solution is only as good as the appropriateness of the turbulence model. In spite of this limitation, the standard turbulence models yield reasonable results for many practical engineering problems. There are several aspects of CFD that are not covered in this chapter— grid generation techniques, numerical algorithms, finite difference and finite volume schemes, stability issues, turbulence modeling, etc. You need to study these topics in order to fully understand both the capabilities and limitations of computational fluid dynamics. In this chapter, we merely scratch the surface of this exciting field. Our goal is to present the fundamentals of CFD from a user’s point of view, providing guidelines about how to generate a grid, how to specify boundary conditions, and how to determine if the computer output is physically meaningful. We begin this section by presenting the differential equations of fluid flow that are to be solved, and then we outline a solution procedure. Subsequent sections of this chapter are devoted to example CFD solutions for laminar flow, turbulent flow, flows with heat transfer, compressible flow, and openchannel flow.
Equations of Motion
For steady laminar flow of a viscous, incompressible, Newtonian fluid without free-surface effects, the equations of motion are the continuity equation →
→
§! V "0
(15–1)
→ → → → 1→ (V ! §)V " # §P$ % n§ 2 V r
(15–2)
and the Navier–Stokes equation
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819 CHAPTER 15
Strictly speaking, Eq. 15–1 is a conservation equation, while Eq. 15–2 is a transport equation that represents transport of linear momentum through→ out the computational domain. In Eqs. 15–1 and 15–2, V is the velocity of the fluid, r is its density, and n is its kinematic viscosity (n " m/r). The lack of free-surface effects enables us to use the modified pressure P$, thereby eliminating the gravity term from Eq. 15–2 (see Chap. 10). Note that Eq. 15–1 is a scalar equation, while Eq. 15–2 is a vector equation. Equations 15–1 and 15–2 apply only to incompressible flows in which we also assume that both r and n are constants. Thus, for three-dimensional flow in Cartesian coordinates, there are four coupled differential equations for four unknowns, u, v, w, and P$ (Fig. 15–2). If the flow were compressible, Eqs. 15–1 and 15–2 would need to be modified appropriately, as discussed in Section 15–5. Liquid flows can almost always be treated as incompressible, and for many gas flows, the gas is at a low enough Mach number that it behaves as a nearly incompressible fluid.
Continuity: ∂v ∂w ∂u + + =0 ∂y ∂z ∂x x-momentum: u
∂u ∂u ∂u +w +v = ∂x ∂y ∂z –
y-momentum:
u
∂v ∂v ∂v +w +v = ∂x ∂y ∂z –
Solution Procedure
To solve Eqs. 15–1 and 15–2 numerically, the following steps are performed. Note that the order of some of the steps (particularly steps 2 through 5) is interchangeable. 1. A computational domain is chosen, and a grid (also called a mesh) is generated; the domain is divided into many small elements called cells. For two-dimensional (2-D) domains, the cells are areas, while for threedimensional (3-D) domains the cells are volumes (Fig. 15–3). You can think of each cell as a tiny control volume in which discretized versions of the conservation equations are solved. Note that we limit our discussion here to cell-centered finite volume CFD codes. The quality of a CFD solution is highly dependent on the quality of the grid. Therefore, you are advised to make sure that your grid is of high quality before proceeding to the next step (Fig. 15–4). 2. Boundary conditions are specified on each edge of the computational domain (2-D flows) or on each face of the domain (3-D flows). 3. The type of fluid (water, air, gasoline, etc.) is specified, along with fluid properties (temperature, density, viscosity, etc.). Many CFD codes Computational domain
∂2v ∂2v ∂2v 1 ∂P' + nQ 2 + 2 + 2 Q ∂x ∂y ∂z r ∂y z-momentum:
u
∂w ∂w ∂w +w +v = ∂x ∂y ∂z –
1 ∂P' ∂2w ∂2w ∂2w + nQ 2 + 2 + 2 Q ∂x ∂y ∂z r ∂z
FIGURE 15–2 The equations of motion to be solved by CFD for the case of steady, incompressible, laminar flow of a Newtonian fluid with constant properties and without free-surface effects. A Cartesian coordinate system is used. There are four equations and four unknowns: u, v, w, and P$.
Computational domain
Cell Cell
Boundaries Boundaries (a)
1 ∂P' ∂2u ∂2u ∂2u + nQ 2 + 2 + 2 Q ∂x ∂y ∂z r ∂x
(b)
FIGURE 15–3 A computational domain is the region in space in which the equations of motion are solved by CFD. A cell is a small subset of the computational domain. Shown are (a) a twodimensional domain and quadrilateral cell, and (b) a three-dimensional domain and hexahedral cell. The boundaries of a 2-D domain are called edges, while those of a 3-D domain are called faces.
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820 FLUID MECHANICS
NOTICE
4.
Do not proceed with CFD calculations until you have generated a highquality grid.
5.
6.
FIGURE 15–4 A quality grid is essential to a quality CFD simulation.
7. FL
FD M
FIGURE 15–5 Global properties of a flow, such as forces and moments on an object, are calculated after a CFD solution has converged. They can also be calculated during the iteration process to monitor convergence.
8.
have built-in property databases for common fluids, making this step relatively painless. Numerical parameters and solution algorithms are selected. These are specific to each CFD code and are not discussed here. The default settings of most modern CFD codes are appropriate for the simple problems discussed in this chapter. Starting values for all flow field variables are specified for each cell. These are initial conditions, which may or may not be correct, but are necessary as a starting point, so that the iteration process may proceed (step 6). We note that for proper unsteady-flow calculations, the initial conditions must be correct. Beginning with the initial guesses, discretized forms of Eqs. 15–1 and 15–2 are solved iteratively, usually at the center of each cell. If one were to put all the terms of Eq. 15–2 on one side of the equation, the solution would be “exact” when the sum of these terms, defined as the residual, is zero for every cell in the domain. In a CFD solution, however, the sum is never identically zero, but (hopefully) decreases with progressive iterations. A residual can be thought of as a measure of how much the solution to a given transport equation deviates from exact, and one monitors the average residual associated with each transport equation to help determine when the solution has converged. Sometimes hundreds or even thousands of iterations are required to converge on a final solution, and the residuals may decrease by several orders of magnitude. Once the solution has converged, flow field variables such as velocity and pressure are plotted and analyzed graphically. Users can also define and analyze additional custom functions that are formed by algebraic combinations of flow field variables. Most commercial CFD codes have built in postprocessors, designed for quick graphical analysis of the flow field. There are also stand-alone postprocessor software packages available for this purpose. Since the graphics output is often displayed in vivid colors, CFD has earned the nickname colorful fluid dynamics. Global properties of the flow field, such as pressure drop, and integral properties, such as forces (lift and drag) and moments acting on a body, are calculated from the converged solution (Fig. 15–5). With most CFD codes, this can also be done “on the fly” as the iterations proceed. In many cases, in fact, it is wise to monitor these quantities along with the residuals during the iteration process; when a solution has converged, the global and integral properties should settle down to constant values as well.
For unsteady flow, a physical time step is specified, appropriate initial conditions are assigned, and an iteration loop is carried out to solve the transport equations to simulate changes in the flow field over this small span of time. Since the changes between time steps are small, a relatively small number of iterations (on the order of tens) is usually required between each time step. Upon convergence of this “inner loop,” the code marches to the next time step. If a flow has a steady-state solution, that solution is often easier to find by marching in time—after enough time has past, the flow field variables settle down to their steady-state values. Most CFD codes take advantage of this fact by internally specifying a pseudo-time step (artificial time) and marching toward a steady-state solution. In such cases, the
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821 CHAPTER 15
pseudo-time step can even be different for different cells in the computational domain and can be tuned appropriately to decrease convergence time. Other “tricks” are often used to reduce computation time, such as multigridding, in which the flow field variables are updated first on a coarse grid so that gross features of the flow are quickly established. That solution is then interpolated to finer and finer grids, the final grid being the one specified by the user (Fig. 15–6). In some commercial CFD codes, several layers of multigridding may occur “behind the scenes” during the iteration process, without user input (or awareness). You can learn more about computational algorithms and other numerical techniques that improve convergence by reading books devoted to computational methods, such as Tannehill, Anderson, and Pletcher (1997).
Additional Equations of Motion
If energy conversion or heat transfer is important in the problem, another transport equation, the energy equation, must also be solved. If temperature differences lead to significant changes in density, an equation of state (such as the ideal-gas law) is used. If buoyancy is important, the effect of temperature on density is reflected in the gravity term (which must then be separated from the modified pressure term in Eq. 15–2). For a given set of boundary conditions, a laminar flow CFD solution approaches an “exact” solution, limited only by the accuracy of the discretization scheme used for the equations of motion, the level of convergence, and the degree to which the grid is resolved. The same would be true of a turbulent flow simulation if the grid could be fine enough to resolve all the unsteady, three-dimensional turbulent eddies. Unfortunately, this kind of direct simulation of turbulent flow is usually not possible for practical engineering applications due to computer limitations. Instead, additional approximations are made in the form of turbulence models so that turbulent flow solutions are possible. The turbulence models generate additional transport equations that model the enhanced mixing and diffusion of turbulence; these additional transport equations must be solved along with those of mass and momentum. Turbulence modeling is discussed in more detail in Section 15–3. Modern CFD codes include options for calculation of particle trajectories, species transport, heat transfer, and turbulence. The codes are easy to use, and solutions can be obtained without knowledge about the equations or their limitations. Herein lies the danger of CFD: When in the hands of someone without knowledge of fluid mechanics, erroneous results are likely to occur (Fig. 15–7). It is critical that users of CFD possess some fundamental knowledge of fluid mechanics so that they can discern whether a CFD solution makes physical sense or not.
FIGURE 15–6 With multigridding, solutions of the equations of motion are obtained on a coarse grid first, followed by successively finer grids. This speeds up convergence.
Grid Generation and Grid Independence
The first step (and arguably the most important step) in a CFD solution is generation of a grid that defines the cells on which flow variables (velocity, pressure, etc.) are calculated throughout the computational domain. Modern commercial CFD codes come with their own grid generators, and thirdparty grid generation programs are also available. The grids used in this chapter are generated with FLUENT’s grid generation package, GAMBIT.
FIGURE 15–7 CFD solutions are easy to obtain, and the graphical outputs can be beautiful; but correct answers depend on correct inputs and knowledge about the flow field.
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822 FLUID MECHANICS y
j=4 3 2 1 i=1
2
3
4
5 6 7 8 x
FIGURE 15–8 Sample structured 2-D grid with nine nodes and eight intervals on the top and bottom edges, and five nodes and four intervals on the left and right edges. Indices i and j are shown. The shaded cell is at (i " 4, j " 3).
Many CFD codes can run with either structured or unstructured grids. A structured grid consists of planar cells with four edges (2-D) or volumetric cells with six faces (3-D). Although the cells can be distorted from rectangular, each cell is numbered according to indices (i, j, k) that do not necessarily correspond to coordinates x, y, and z. An illustration of a 2-D structured grid is shown in Fig. 15–8. To construct this grid, nine nodes are specified on the top and bottom edges; these nodes correspond to eight intervals along these edges. Similarly, five nodes are specified on the left and right edges, corresponding to four intervals along these edges. The intervals correspond to i " 1 through 8 and j " 1 through 4, and are numbered and marked in Fig. 15–8. An internal grid is then generated by connecting nodes one-for-one across the domain such that rows ( j " constant) and columns (i " constant) are clearly defined, even though the cells themselves may be distorted (not necessarily rectangular). In a 2-D structured grid, each cell is uniquely specified by an index pair (i, j). For example, the shaded cell in Fig. 15–8 is at (i " 4, j " 3). You should be aware that some CFD codes number nodes rather than intervals. An unstructured grid consists of cells of various shapes, but typically triangles or quadrilaterals (2-D) and tetrahedrons or hexahedrons (3-D) are used. Two unstructured grids for the same domain as that of Fig. 15–8 are generated, using the same interval distribution on the edges; these grids are shown in Fig. 15–9. Unlike the structured grid, one cannot uniquely identify cells in the unstructured grid by indices i and j; instead, cells are numbered in some other fashion internally in the CFD code. For complex geometries, an unstructured grid is usually much easier for the user of the grid generation code to create. However, there are some advantages to structured grids. For example, some (usually older) CFD codes are written specifically for structured grids; these codes converge more rapidly, and often more accurately, by utilizing the index feature of structured grids. For modern general-purpose CFD codes that can handle both structured and unstructured grids, however, this is no longer an issue. More importantly, fewer cells are usually generated with a structured grid than with an unstructured grid. In Fig. 15–8, for example, the structured grid has 8 & 4 " 32 cells, while the unstructured triangular grid of Fig. 15–9a has 76 cells, and the unstructured quadrilateral grid has 38 cells, even though the identical node distribution is applied at the edges in all three cases. In boundy
FIGURE 15–9 Sample 2-D unstructured grids with nine nodes and eight intervals on the top and bottom edges, and five nodes and four intervals on the left and right edges. These grids use the same node distribution as that of Fig. 15–8: (a) unstructured triangular grid, and (b) unstructured quadrilateral grid. The shaded cell in (a) is moderately skewed.
y
Unstructured triangular grid (a)
Unstructured quadrilateral grid x
(b)
x
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823 CHAPTER 15
ary layers, where flow variables change rapidly normal to the wall and highly resolved grids are required close to the wall, structured grids enable much finer resolution than do unstructured grids for the same number of cells. This can be seen by comparing the grids of Figs. 15–8 and 15–9 near the far right edge. The cells of the structured grid are thin and tightly packed near the right edge, while those of the unstructured grids are not. We must emphasize that regardless of the type of grid you choose (structured or unstructured, quadrilateral or triangular, etc.), it is the quality of the grid that is most critical for reliable CFD solutions. In particular, you must always be careful that individual cells are not highly skewed, as this can lead to convergence difficulties and inaccuracies in the numerical solution. The shaded cell in Fig. 15–9a is an example of a cell with moderately high skewness, defined as the departure from symmetry. There are various kinds of skewness, for both two- and three-dimensional cells. Three-dimensional cell skewness is beyond the scope of the present textbook—the type of skewness most appropriate for two-dimensional cells is equiangle skewness, defined as Equiangle skewness:
u max # u equal u equal # u min Q EAS " MAX a , b 180( # u equal u equal
(15–3)
where umin and umax are the minimum and maximum angles (in degrees) between any two edges of the cell, and uequal is the angle between any two edges of an ideal equilateral cell with the same number of edges. For triangular cells uequal " 60° and for quadrilateral cells uequal " 90°. You can show by Eq. 15–3 that 0 ' QEAS ' 1 for any 2-D cell. By definition, an equilateral triangle has zero skewness. In the same way, a square or rectangle has zero skewness. A grossly distorted triangular or quadrilateral element may have unacceptably high skewness (Fig. 15–10). Some grid generation codes use numerical schemes to smooth the grid so as to minimize skewness. Other factors affect the quality of the grid as well. For example, abrupt changes in cell size can lead to numerical or convergence difficulties in the CFD code. Also, cells with a very large aspect ratio can sometimes cause problems. While you can often minimize the cell count by using a structured grid instead of an unstructured grid, a structured grid is not always the best choice, depending on the shape of the computational domain. You must always be cognizant of grid quality. Keep in mind that a high-quality unstructured grid is better than a poor-quality structured grid. An example is shown in Fig. 15–11 for the case of a computational domain with a small acute angle at the upper-right corner. For this example we have adjusted the
(a)
(b)
(c)
(d)
(a) Triangular cells
Zero skewness
High skewness
(b) Quadrilateral cells
Zero skewness
High skewness
FIGURE 15–10 Skewness is shown in two dimensions: (a) an equilateral triangle has zero skewness, but a highly distorted triangle has high skewness. (b) Similarly, a rectangle has zero skewness, but a highly distorted quadrilateral cell has high skewness.
FIGURE 15–11 Comparison of four 2-D grids for a highly distorted computational domain: (a) structured 8 & 8 grid with 64 cells and (QEAS)max " 0.83, (b) unstructured triangular grid with 70 cells and (QEAS)max " 0.76, (c) unstructured quadrilateral grid with 67 cells and (QEAS)max " 0.87, and (d) hybrid grid with 62 cells and (QEAS)max " 0.76.
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824 FLUID MECHANICS
FIGURE 15–12 Examples of structured grids generated for multiblock CFD analysis: (a) a simple 2-D computational domain composed of rectangular four-sided blocks, and (b) a more complicated 2-D domain with curved surfaces, but again composed of four-sided blocks and quadrilateral cells. The number of i- and j-intervals is shown in parentheses for each block. There are, of course, acceptable alternative ways to divide these computational domains into blocks.
Block 2 (10 & 8) Block 7 (3 & 5) Block 1 (12 & 8) Block 3 (10 & 8) Block 4 (10 & 5)
Block 5 (6 & 5)
Block 6 (3 & 5)
FIGURE 15–13 The multiblock grid of Fig. 15–12a modified for a CFD code that can handle only elementary blocks.
Block 3 (5 & 8) Block 2 (5 & 16) Block 4 (3 & 5) Block 1 (12 & 8) Block 2 (10 & 21) Block 3 (9 & 5) (a)
Block 4 (5 & 16)
Block 6 Block 5 (8 & 16) (5 & 8)
Block 1 (12 & 8)
(b)
node distribution so that the grid in any case contains between 60 and 70 cells for direct comparison. The structured grid (Fig. 15–11a) has 8 & 8 " 64 cells; but even after smoothing, the maximum equiangle skewness is 0.83—cells near the upper right corner are highly skewed. The unstructured triangular grid (Fig. 15–11b) has 70 cells, but the maximum skewness is reduced to 0.76. More importantly, the overall skewness is lower throughout the entire computational domain. The unstructured quad grid (Fig. 15–11c) has 67 cells. Although the overall skewness is better than that of the structured mesh, the maximum skewness is 0.87—higher than that of the structured mesh. The hybrid grid shown in Fig. 15–11d is discussed shortly. Situations arise in which a structured grid is preferred (e.g., the CFD code requires structured grids, boundary layer zones need high resolution, or the simulation is pushing the limits of available computer memory). Generation of a structured grid is straightforward for geometries with straight edges. All we need to do is divide the computational domain into four-sided (2-D) or six-sided (3-D) blocks or zones. Inside each block, we generate a structured grid (Fig. 15–12a). Such an analysis is called multiblock analysis. For more complicated geometries with curved surfaces, we need to determine how the computational domain can be divided into individual blocks that may or may not have flat edges (2-D) or faces (3-D). A two-dimensional example involving circular arcs is shown in Fig. 15–12b. Most CFD codes require that the nodes match on the common edges and faces between blocks. Many commercial CFD codes allow you to split the edges or faces of a block and assign different boundary conditions to each segment of the edge or face. In Fig. 15–12a for example, the left edge of block 2 is split about two-thirds of the way up to accommodate the junction with block 1. The lower segment of this edge is a wall, and the upper segment of this edge is an interior edge. (These and other boundary conditions are discussed shortly.) A similar situation occurs on the right edge of block 2 and on the top edge of block 3. Some CFD codes accept only elementary blocks, namely, blocks whose edges or faces cannot be split. For example, the fourblock grid of Fig. 15–12a requires seven elementary blocks under this limitation (Fig. 15–13). The total number of cells is the same, which you can
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825 CHAPTER 15
verify. Finally, for CFD codes that allow blocks with split edges or faces, we can sometimes combine two or more blocks into one. For example, it is left as an exercise to show how the structured grid of Fig. 5-11b can be simplified to just three nonelementary blocks. When developing the block topology with complicated geometries as in Fig. 15–12b, the goal is to create blocks in such a way that no cells in the grid are highly skewed. In addition, cell size should not change abruptly in any direction, and the blocking topology should lend itself to clustering cells near solid walls so that boundary layers can be resolved. With practice you can master the art of creating sophisticated multiblock structured grids. Multiblock grids are necessary for structured grids of complex geometry. They may also be used with unstructured grids, but are not necessary since the cells can accommodate complex geometries. Finally, a hybrid grid is one that combines regions or blocks of structured and unstructured grids. For example, you can mate a structured grid block close to a wall with an unstructured grid block outside of the region of influence of the boundary layer. A hybrid grid is often used to enable high resolution near a wall without requiring high resolution away from the wall (Fig. 15–14). When generating any type of grid (structured, unstructured, or hybrid), you must always be careful that individual cells are not highly skewed. For example, none of the cells in Fig. 15–14 has any significant skewness. Another example of a hybrid grid is shown in Fig. 15–11d. Here we have split the computational domain into two blocks. The foursided block on the left is meshed with a structured grid, while the threesided block on the right is meshed with an unstructured triangular grid. The maximum skewness is 0.76, the same as that of the unstructured triangular grid of Fig. 15–11b, but the total number of cells is reduced from 70 to 62. Computational domains with very small angles like the one shown in Fig. 15–11 are difficult to mesh at the sharp corner, regardless of the type of cells used. One way to avoid large values of skewness at a sharp corner is to simply chop off or round off the sharp corner. This can be done very close to the corner so that the geometric modification is imperceptible from an overall view and has little if any effect on the flow, yet greatly improves the performance of the CFD code by reducing the skewness. For example, the troublesome sharp corner of the computational domain of Fig. 15–11 is chopped off and replotted in Fig. 15–15. Through use of multiblocking and hybrid grids, the grid shown in Fig. 15–15 has 62 cells and a maximum skewness of only 0.53—a vast improvement over any of the grids in Fig. 15–11. Generation of a good grid is often tedious and time consuming; engineers who use CFD on a regular basis will agree that grid generation usually takes more of their time than does the CFD solution itself (engineer’s time, not
(a)
(b)
Structured Unstructured
Structured
FIGURE 15–14 Sample two-dimensional hybrid grid near a curved surface; two structured regions and one unstructured region are labeled.
FIGURE 15–15 Hybrid grid for the computational domain of Fig. 15–11 with the sharp corner chopped off: (a) overall view—the grid contains 62 cells with (QEAS)max " 0.53, (b) magnified view of the chopped off corner.
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826 FLUID MECHANICS
FIGURE 15–16 Time spent generating a good grid is time well spent.
Boundary Conditions
Wall Inlet
Computational domain
CPU time). However, time spent generating a good grid is time well spent, since the CFD results will be more reliable and may converge more rapidly (Fig. 15–16). A high-quality grid is critical to an accurate CFD solution; a poorly resolved or low-quality grid may even lead to an incorrect solution. It is important, therefore, for users of CFD to test if their solution is grid independent. The standard method to test for grid independence is to increase the resolution (by a factor of 2 in all directions if feasible) and repeat the simulation. If the results do not change appreciably, the original grid is probably adequate. If, on the other hand, there are significant differences between the two solutions, the original grid is likely of inadequate resolution. In such a case, an even finer grid should be tried until the grid is adequately resolved. This method of testing for grid independence is time consuming, and unfortunately, not always feasible, especially for large engineering problems in which the solution pushes computer resources to their limits. In a 2-D simulation, if one doubles the number of intervals on each edge, the number of cells increases by a factor of 22 " 4; the required computation time for the CFD solution also increases by approximately a factor of 4. For three-dimensional flows, doubling the number of intervals in each direction increases the cell count by a factor of 23 " 8. You can see how grid independence studies can easily get beyond the range of a computer’s memory capacity and/or CPU availability. If you cannot double the number of intervals because of computer limitations, a good rule of thumb is that you should increase the number of intervals by at least 20 percent in all directions to test for grid independence. On a final note about grid generation, the trend in CFD today is automated grid generation, coupled with automated grid refinement based on error estimates. Yet even in the face of these emerging trends, it is critical that you understand how the grid impacts the CFD solution.
Outlet
Wall
FIGURE 15–17 Boundary conditions must be carefully applied at all boundaries of the computational domain. Appropriate boundary conditions are required in order to obtain an accurate CFD solution.
While the equations of motion, the computational domain, and even the grid may be the same for two CFD calculations, the type of flow that is modeled is determined by the imposed boundary conditions. Appropriate boundary conditions are required in order to obtain an accurate CFD solution (Fig. 15–17). There are several types of boundary conditions available; the most relevant ones are listed and briefly described in the following. The names are those used by FLUENT; other CFD codes may use somewhat different terminology, and the details of their boundary conditions may differ. In the descriptions given, the words face or plane are used, implying three-dimensional flow. For a two-dimensional flow, the word edge or line should be substituted for face or plane.
Wall Boundary Conditions
The simplest boundary condition is that of a wall. Since fluid cannot pass through a wall, the normal component of velocity is set to zero relative to the wall along a face on which the wall boundary condition is prescribed. In addition, because of the no-slip condition, we usually set the tangential component of velocity at a stationary wall to zero as well. In Fig. 15–17, for example, the upper and lower edges of this simple domain are specified as
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827 CHAPTER 15
wall boundary conditions with no slip. If the energy equation is being solved, either wall temperature or wall heat flux must also be specified (but not both; see Section 15–4). If a turbulence model is being used, turbulence transport equations are solved, and wall roughness may need to be specified, since turbulent boundary layers are influenced greatly by the roughness of the wall. In addition, users must choose among various kinds of turbulence wall treatments (wall functions, etc.). These turbulence options are beyond the scope of the present text (see Wilcox, 1998); fortunately the default options of most modern CFD codes are sufficient for many applications involving turbulent flow. Moving walls and walls with specified shear stresses can also be simulated in many CFD codes. There are situations where we desire to let the fluid slip along the wall (we call this an “inviscid wall”). For example, we can specify a zero-shear-stress wall boundary condition along the free surface of a swimming pool or hot tub when simulating such a flow (Fig. 15–18). Note that with this simplification, the fluid is allowed to “slip” along the surface, since the viscous shear stress caused by the air above it is negligibly small (Chap. 9). When making this approximation, however, surface waves and their associated pressure fluctuations cannot be taken into account.
Inflow/Outflow Boundary Conditions
There are several options at the boundaries through which fluid enters the computational domain (inflow) or leaves the domain (outflow). They are generally categorized as either velocity-specified conditions or pressurespecified conditions. At a velocity inlet, we specify the velocity of the incoming flow along the inlet face. If energy and/or turbulence equations are being solved, the temperature and/or turbulence properties of the incoming flow need to be specified as well. At a pressure inlet, we specify the total pressure along the inlet face (for example, flow coming into the computational domain from a pressurized tank of known pressure or from the far field where the ambient pressure is known). At a pressure outlet, fluid flows out of the computational domain. We specify the static pressure along the outlet face; in many cases this is atmospheric pressure (zero gage pressure). For example, the pressure is atmospheric at the outlet of a subsonic exhaust pipe open to ambient air (Fig. 15–19). Flow properties, such as temperature, and turbulence properties are also specified at pressure inlets and pressure outlets. For the latter case, however, these properties are not used unless the solution demands reverse flow across the outlet. Reverse flow at a pressure outlet is usually an indication that the computational domain is not large enough. If reverse flow warnings persist as the CFD solution iterates, the computational domain should be extended. Pressure is not specified at a velocity inlet, as this would lead to mathematical overspecification, since pressure and velocity are coupled in the equations of motion. Rather, pressure at a velocity inlet adjusts itself to match the rest of the flow field. In similar fashion, velocity is not specified at a pressure inlet or outlet, as this would also lead to mathematical overspecification. Rather, velocity at a pressure-specified boundary condition adjusts itself to match the rest of the flow field (Fig. 15–20).
The free surface is approximated as a wall boundary condition with slip (zero shear stress). Velocity inlet
Computational domain
Vin
Standard no-slip wall boundary condition Pressure outlet
Pout
FIGURE 15–18 The standard wall boundary condition is imposed on stationary solid boundaries, where we also impose either a wall temperature or a wall heat flux. The shear stress along the wall can be set to zero to simulate the free surface of a liquid, as shown here for the case of a swimming pool. There is slip along this “wall” that simulates the free surface (in contact with air).
Pout out = Patm atm
Pressure outlet
FIGURE 15–19 When modeling an incompressible flow field, with the outlet of a pipe or duct exposed to ambient air, the proper boundary condition is a pressure outlet with Pout " Patm. Shown here is the tail pipe of an automobile. Photo by Po-Ya Abel Chuang. Used by permission.
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828 FLUID MECHANICS
FIGURE 15–20 At a pressure inlet or a pressure outlet, we specify the pressure on the face, but we cannot specify the velocity through the face. As the CFD solution converges, the velocity adjusts itself such that the prescribed pressure boundary conditions are satisfied.
u
Outflow boundary x
FIGURE 15–21 At an outflow boundary condition, the gradient or slope of velocity normal to the outflow face is zero, as illustrated here for u as a function of x along a horizontal line. Note that neither pressure nor velocity are specified at an outflow boundary.
Pin
Pout
Inlet velocity calculated, not specified
Computational domain
Pressure inlet; Pin specified
Outlet velocity calculated, not specified
Pressure outlet; Pout specified
Another option at an outlet of the computational domain is the outflow boundary condition. At an outflow boundary, no flow properties are specified; instead, flow properties such as velocity, turbulence quantities, and temperature are forced to have zero gradients normal to the outflow face (Fig. 15–21). For example, if a duct is sufficiently long so that the flow is fully developed at the outlet, the outflow boundary condition would be appropriate, since velocity does not change in the direction normal to the outlet face. Note that the flow direction is not constrained to be perpendicular to the outflow boundary, as also illustrated in Fig. 15–21. If the flow is still developing, but the pressure at the outlet is known, a pressure outlet boundary condition would be more appropriate than an outflow boundary condition. The outflow boundary condition is often preferred over the pressure outlet in rotating flows since the swirling motion leads to radial pressure gradients that are not easily handled by a pressure outlet. A common situation in a simple CFD application is to specify one or more velocity inlets along portions of the boundary of the computational domain, and one or more pressure outlets or outflows at other portions of the boundary, with walls defining the geometry of the rest of the computational domain. For example, in our swimming pool (Fig. 15–18), we set the left-most face of the computational domain as a velocity inlet and the bottom-most face as a pressure outlet. The rest of the faces are walls, with the free surface modeled as a wall with zero shear stress. Finally, for compressible flow simulations, the inlet and outlet boundary conditions are further complicated by introduction of Riemann invariants and characteristic variables related to incoming and outgoing waves, discussion of which is beyond the scope of the present text. Fortunately, many CFD codes have a pressure far field boundary condition for compressible flows. This boundary condition is used to specify the Mach number, pressure, and temperature at an inlet. The same boundary condition can be applied at an outlet; when flow exits the computational domain, flow variables at the outlet are extrapolated from the interior of the domain. Again you must ensure that there is no reverse flow at an outlet.
Miscellaneous Boundary Conditions
Some boundaries of a computational domain are neither walls nor inlets or outlets, but rather enforce some kind of symmetry or periodicity. For example, the periodic boundary condition is useful when the geometry involves repetition. Flow field variables along one face of a periodic boundary are
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numerically linked to a second face of identical shape (and in most CFD codes, also identical face mesh). Thus, flow leaving (crossing) the first periodic boundary can be imagined as entering (crossing) the second periodic boundary with identical properties (velocity, pressure, temperature, etc.). Periodic boundary conditions always occur in pairs and are useful for flows with repetitive geometries, such as flow between the blades of a turbomachine or through an array of heat exchanger tubes (Fig. 15–22). The periodic boundary condition enables us to work with a computational domain that is much smaller than the full flow field, thereby conserving computer resources. In Fig. 15–22, you can imagine an infinite number of repeated domains (dashed lines) above and below the actual computational domain (the light blue shaded region). Periodic boundary conditions must be specified as either translational (periodicity applied to two parallel faces, as in Fig. 15–22) or rotational (periodicity applied to two radially oriented faces). The region of flow between two neighboring blades of a fan (a flow passage) is an example of a rotationally periodic domain (see Fig. 15–56). The symmetry boundary condition forces flow field variables to be mirrorimaged across a symmetry plane. Mathematically, gradients of most flow field variables in the direction normal to the symmetry plane are set to zero across the plane of symmetry, although some variables are specified as even functions and some as odd functions across a symmetry boundary condition. For physical flows with one or more symmetry planes, this boundary condition enables us to model a portion of the physical flow domain, thereby conserving computer resources. The symmetry boundary differs from the periodic boundary in that no “partner” boundary is required for the symmetry case. In addition, fluid can flow parallel to a symmetry boundary, but not through a symmetry boundary, whereas flow can cross a periodic boundary. Consider, for example, flow across an array of heat exchanger tubes (Fig. 15–22). If we assume that no flow crosses the periodic boundaries of that computational domain, we can use symmetry boundary conditions instead. Alert readers will notice that we can even cut the size of the computational domain in half by wise choice of symmetry planes (Fig. 15–23). For axisymmetric flows, the axis boundary condition is applied to a straight edge that represents the axis of symmetry (Fig. 15–24a). Fluid can flow parallel to the axis, but cannot flow through the axis. The axisymmetric option enables us to solve the flow in only two dimensions, as sketched in Fig. 15–24b. The computational domain is simply a rectangle in the xyplane; you can imagine rotating this plane about the x-axis to generate the axisymmetry. In the case of swirling axisymmetric flows, fluid can also flow tangentially in a circular path around the axis of symmetry. Swirling axisymmetric flows are sometimes called rotationally symmetric.
Internal Boundary Conditions
The final classification of boundary conditions is imposed on faces or edges that do not define a boundary of the computational domain, but rather exist inside the domain. When an interior boundary condition is specified on a face, flow crosses through the face without any user-forced changes, just as it would cross from one interior cell to another (Fig. 15–25). This boundary condition is necessary for situations in which the computational domain is divided into separate blocks or zones, and enables
Periodic
In
Computational domain
Out
Periodic
FIGURE 15–22 The periodic boundary condition is imposed on two identical faces. Whatever happens at one of the faces must also happen at its periodic partner face, as illustrated by the velocity vectors crossing the periodic faces.
Symmetry In
Computational domain
Out
Symmetry
FIGURE 15–23 The symmetry boundary condition is imposed on a face so that the flow across that face is a mirror image of the calculated flow. We sketch imaginary domains (dashed lines) above and below the computational domain (the light blue shaded region) in which the velocity vectors are mirror images of those in the computational domain. In this heat exchanger example, the left face of the domain is a velocity inlet, the right face is a pressure outlet or outflow outlet, the cylinders are walls, and both the top and bottom faces are symmetry planes.
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830 FLUID MECHANICS y
FIGURE 15–24 The axis boundary condition is applied to the axis of symmetry (here the x-axis) in an axisymmetric flow, since there is rotational symmetry about that axis. (a) A slice defining the xy- or ru-plane is shown, and the velocity components can be either (u, v) or (ur, uu). (b) The computational domain (light blue shaded region) for this problem is reduced to a plane in two dimensions (x and y). In many CFD codes, x and y are used as axisymmetric coordinates, with y being understood as the distance from the x-axis.
Out P + ∆P P In
Fan
Interior
FIGURE 15–25 The fan boundary condition imposes an abrupt change in pressure across the fan face to simulate an axial-flow fan in a duct. When the specified pressure rise is zero, the fan boundary condition degenerates to an interior boundary condition.
y
v →
V uu
Computational domain
ur
→
r
In
u
V
v
u
x Rotational symmetry
y
Axis
Axis
Out
u x
Wall
Axisymmetric body x (a)
(b)
communication between blocks. We have found this boundary condition to be useful for postprocessing as well, since a predefined face is present in the flow field, on whose surface one can plot velocity vectors, pressure contours, etc. In more sophisticated CFD applications in which there is a sliding or rotating mesh, the interface between the two blocks is called upon to smoothly transfer information from one block to another. The fan boundary condition is specified on a plane across which a sudden pressure increase (or decrease) is to be assigned. This boundary condition is similar to an interior boundary condition except for the forced pressure rise. The CFD code does not solve the detailed, unsteady flow field through individual fan blades, but simply models the plane as an infinitesimally thin fan that changes the pressure across the plane. The fan boundary condition is useful, for example, as a simple model of a fan inside a duct (Fig. 15–25), a ceiling fan in a room, or the propeller or jet engine that provides thrust to an airplane. If the pressure rise across the fan is specified as zero, this boundary condition behaves the same as an interior boundary condition.
Practice Makes Perfect
The best way to learn computational fluid dynamics is through examples and practice. You are encouraged to experiment with various grids, boundary conditions, numerical parameters, etc., in order to get a feel for CFD. Before tackling a complicated problem, it is best to solve simpler problems, especially ones for which analytical or empirical solutions are known (for comparison and verification). In the following sections, we solve several example problems of general engineering interest to illustrate many of the capabilities and limitations of CFD. We start with laminar flows, and then provide some introductory turbulent flow examples. Finally we provide examples of flows with heat transfer, compressible flows, and liquid flows with free surfaces. Color images of the results are available on the book’s website, including some animations.
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15–2
■
LAMINAR CFD CALCULATIONS
Computational fluid dynamics does an excellent job at computing incompressible, steady or unsteady, laminar flow, provided that the grid is well resolved and the boundary conditions are properly specified. We show several simple examples of laminar flow solutions, paying particular attention to grid resolution and appropriate application of boundary conditions. In all examples in this section, the flows are incompressible and two-dimensional (or axisymmetric).
Pipe Flow Entrance Region at Re ! 500
Consider flow of room-temperature water inside a smooth round pipe of length L " 40.0 cm and diameter D " 1.00 cm. We assume that the water enters at a uniform speed equal to V " 0.05024 m/s. The kinematic viscosity of the water is n " 1.005 & 10#6 m2/s, producing a Reynolds number of Re " VD/n " 500. We assume incompressible, steady, laminar flow. We are interested in the entrance region in which the flow gradually becomes fully developed. Because of the axisymmetry, we set up a computational domain that is a two-dimensional slice from the axis to the wall, rather than a threedimensional cylindrical volume (Fig. 15–26). We generate six structured grids for this computational domain: very coarse (40 intervals in the axial direction & 8 intervals in the radial direction), coarse (80 & 16), medium (160 & 32), fine (320 & 64), very fine (640 & 128), and ultrafine (1280 & 256). (Note that the number of intervals is doubled in both directions for each successive grid; the number of computational cells increases by a factor of 4 for each grid.) In all cases the nodes are evenly distributed axially, but are concentrated near the wall radially, since we expect larger velocity gradients near the pipe wall. Close-up views of the first three of these grids are shown in Fig. 15–27.
Velocity inlet
Wall
Pressure outlet
Computational domain
r D
Axis
x
L V
FIGURE 15–26 Because of axisymmetry about the x-axis, flow through a round pipe can be solved computationally with a twodimensional slice through the pipe from r " 0 to D/2. The computational domain is the light blue shaded region, and the drawing is not to scale. Boundary conditions are indicated.
(a)
(b)
(c)
FIGURE 15–27 Portions of the three coarsest structured grids generated for the laminar pipe flow example: (a) very coarse (40 & 8), (b) coarse (80 & 16), and (c) medium (160 & 32). The number of computational cells is 320, 1280, and 5120, respectively. In each view, the pipe wall is at the top and the pipe axis is at the bottom, as in Fig. 15–26.
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832 FLUID MECHANICS 100 Continuity
10–2 10–4
x-momentum
10–6 10–8
y-momentum
10–10 10–12 10–14 10–16 0
200 400 Iteration number
600
FIGURE 15–28 Decay of the residuals with iteration number for the very coarse grid laminar pipe flow solution (double precision arithmetic).
We run the CFD program FLUENT in double precision for all six cases. (Double precision arithmetic is not always necessary for engineering calculations—we use it here to obtain the best possible precision in our comparisons.) Since the flow is laminar, incompressible, and axisymmetric, only three transport equations are solved—continuity, x-momentum, and ymomentum. Note that coordinate y is used in the CFD code instead of r as the distance from the axis of rotation (Fig. 15–24). The CFD code is run until convergence (all the residuals level off). Recall that a residual is a measure of how much the solution to a given transport equation deviates from exact; the lower the residual, the better the convergence. For the very coarse grid case, this occurs in about 500 iterations, and the residuals level off to less than 10#12 (relative to their initial values). The decay of the residuals is plotted in Fig. 15–28 for the very coarse case. Note that for more complicated flow problems with finer grids, you cannot always expect such low residuals; in some CFD solutions, the residuals level off at much higher values, like 10#3. We define P1 as the average pressure at an axial location one pipe diameter downstream of the inlet. Similarly we define P20 at 20 pipe diameters. The average axial pressure drop from 1 to 20 diameters is thus )P " P1 # P20, and is equal to 4.404 Pa (to four significant digits of precision) for the very coarse grid case. Centerline pressure and axial velocity are plotted in Fig. 15–29a as functions of downstream distance. The solution appears to be physically reasonable. We see the increase of centerline axial velocity to conserve mass as the boundary layer on the pipe wall grows downstream. We see a sharp drop in pressure near the pipe entrance where viscous shear stresses on the pipe wall are highest. The pressure drop approaches linear toward the end of the entrance region where the flow is nearly fully developed, as expected. Finally, we compare in Fig. 15–29b the axial velocity profile at the end of the pipe to the known analytical solution for fully developed laminar pipe flow (see Chap. 8). The agreement is excellent, especially considering that there are only eight intervals in the radial direction. Is this CFD solution grid independent? To find out, we repeat the calculations using the coarse, medium, fine, very fine, and ultrafine grids. The
8
2
7
CFD
uCL/V
Pgage, Pa
6
FIGURE 15–29 CFD results for the very coarse grid laminar pipe flow simulation: (a) development of centerline pressure and centerline axial velocity with downstream distance, and (b) axial velocity profile at pipe outlet compared to analytical prediction.
0.5 0.4
1.5
5 4
1
3 pgage P
2
0.5
uCL V
r D
0.3 Analytical 0.2 0.1
1 0
0 0
10
20 x/D (a)
30
40
0 0
0.5
1 u/V (b)
1.5
2
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833 CHAPTER 15
TA B L E 1 5 – 1 Pressure drop from x/D " 1 to 20 for the various grid resolution cases in the entrance flow region of axisymmetric pipe flow Very coarse Coarse Medium Fine Very fine Ultrafine
Number of Cells
)P, Pa
320 1,280 5,120 20,480 81,920 327,680
4.404 3.983 3.998 4.016 4.033 4.035
convergence of the residuals is qualitatively similar to that of Fig. 15–28 for all cases, but CPU time increases significantly as grid resolution improves, and the levels of the final residuals are not as low as those of the coarse resolution case. The number of iterations required until convergence also increases with improved grid resolution. The pressure drop from x/D " 1 to 20 is listed in Table 15–1 for all six cases. )P is also plotted as a function of number of cells in Fig. 15–30. We see that even the very coarse grid does a reasonable job at predicting )P. The difference in pressure drop from the very coarse grid to the ultrafine grid is less than 10 percent. Thus, the very coarse grid may be adequate for some engineering calculations. If greater precision is needed, however, we must use a finer grid. We see grid independence to three significant digits by the very fine case. The change in )P from the very fine grid to the ultrafine grid is less than 0.07 percent—a grid as finely resolved as the ultrafine grid is unnecessary in any practical engineering analysis. The most significant differences between the six cases occur very close to the pipe entrance, where pressure gradients and velocity gradients are largest. In fact, there is a singularity at the inlet, where the axial velocity changes suddenly from V to zero at the wall because of the no-slip condition. We plot in Fig. 15–31 contour plots of normalized axial velocity, u/V near the pipe entrance. We see that although global properties of the flow field (like overall pressure drop) vary by only a few percent as the grid is refined, details of the flow field (like the velocity contours shown here) change considerably with grid resolution. You can see that as the grid is continually refined, the axial velocity contour shapes become smoother and more well defined. The greatest differences in the contour shapes occur near the pipe wall.
Flow around a Circular Cylinder at Re ! 150
To illustrate that reliable CFD results require correct problem formulation, consider the seemingly simple problem of steady, incompressible, twodimensional flow over a circular cylinder of diameter D " 2.0 cm (Fig. 15–32). The two-dimensional computational domain used for this simulation
4.5 4.4 4.3 ∆P, Pa
Case
4.2 4.1 4 3.9 3.8 102
103 104 105 Number of cells
106
FIGURE 15–30 Calculated pressure drop from x/D " 1 to 20 in the entrance flow region of axisymmetric pipe flow as a function of number of cells.
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834 FLUID MECHANICS
r 1.1
1.2
1.3
1.4
x (a)
r 1.1
1.2
1.3
1.4
x (b)
r 1.1
1.3
1.2
1.4
x
FIGURE 15–31 Normalized axial velocity contours (u/V) for the laminar pipe flow example. Shown is a close-up view of the entrance region of the pipe for each of the first four grids: (a) very coarse (40 & 8), (b) coarse (80 & 16), (c) medium (160 & 32), and (d) fine (320 & 64).
y V
D
Cylinder
x
FIGURE 15–32 Flow of fluid at free-stream speed V over a two-dimensional circular cylinder of diameter D.
(c)
r 1.1
1.2
1.3
1.4
x (d)
is sketched in Fig. 15–33. Only the upper half of the flow field is solved, due to symmetry along the bottom edge of the computational domain; a symmetry boundary condition is specified along this edge to ensure that no flow crosses the plane of symmetry. With this boundary condition imposed, the required computational domain size is reduced by a factor of 2. A stationary, no-slip wall boundary condition is applied at the cylinder surface. The left half of the far field outer edge of the domain has a velocity inlet boundary condition, on which is specified the velocity components u " V and v " 0. A pressure outlet boundary condition is specified along the right half of the outer edge of the domain. (The gage pressure there is set to zero, but since the velocity field in an incompressible CFD code depends only on pressure differences, not absolute value of pressure, the value of pressure specified for the pressure outlet boundary condition is irrelevant.) Three two-dimensional structured grids are generated for comparison: coarse (30 radial intervals & 60 intervals along the cylinder surface " 1800 cells), medium (60 & 120 " 7200 cells), and fine (120 & 240 " 28,800 cells), as seen in Fig. 15–34. Note that only a small portion of the computational domain is shown here; the full domain extends 15 cylinder diameters
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835 CHAPTER 15
Far field inflow (velocity inlet)
V
Far field outflow (pressure outlet)
Computational domain
Cylinder surface (wall) Symmetry line (symmetry)
y a
x 0.3 m
0.02 m
outward from the origin, and the cells get progressively larger further away from the cylinder. We apply a free-stream flow of air at a temperature of 25°C, at standard atmospheric pressure, and at velocity V " 0.1096 m/s from left to right around this circular cylinder. The Reynolds number of the flow, based on cylinder diameter (D " 2.0 cm), is thus Re " rVD/m " 150. Experiments at this Reynolds number reveal that the boundary layer is laminar and separates almost 10° before the top of the cylinder, at a ! 82° from the front stagnation point. The wake also remains laminar. Experimentally measured values of drag coefficient at this Reynolds number show much discrepancy in the literature; the range is from CD ! 1.1 to 1.4, and the differences are most likely due to the quality of the free-stream and three-dimensional effects (oblique vortex shedding, etc.). (Recall that CD " 2FD /rV 2A, where A is the frontal area of the cylinder, and A " D times the span of the cylinder, taken as unit length in a two-dimensional CFD calculation.) CFD solutions are obtained for each of the three grids shown in Fig. 15–34, assuming steady laminar flow. All three cases converge without problems, but the results do not necessarily agree with physical intuition or with experimental data. Streamlines are shown in Fig. 15–35 for the three grid resolutions. In all cases, the image is mirrored about the symmetry line so that even though only the top half of the flow field is solved, the full flow field is displayed. For the coarse resolution case (Fig. 15–35a), the boundary layer separates at a " 120°, well past the top of the cylinder, and CD is 1.00. The boundary layer is not well enough resolved to yield the proper boundary layer separation point, and the drag is somewhat smaller than it should be. Two large counter-rotating separation bubbles are seen in the wake; they stretch several cylinder diameters downstream. For the medium resolution case (Fig. 15–35b), the flow field is significantly different. The boundary layer separates a little further upstream at a " 110°, which is more in line with the experimental results, but CD has decreased to about 0.982—further away from the experimental value. The separation bubbles in the cylinder’s wake have grown much longer than those of the coarse grid case. Does refining the grid even further improve the numerical results? Figure 15–35c shows streamlines for the fine resolution case. The results look qualitatively similar to those of the medium resolution case, with a " 109°, but the drag coefficient is even smaller (CD " 0.977), and the separation bubbles are even
FIGURE 15–33 Computational domain (light blue shaded region) used to simulate steady two-dimensional flow over a circular cylinder (not to scale). It is assumed that the flow is symmetric about the x-axis. Applied boundary conditions are shown for each edge in parentheses. We also define a, the angle measured along the cylinder surface from the front stagnation point.
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836 FLUID MECHANICS
(a)
(b)
FIGURE 15–34 Structured two-dimensional grids around the upper half of a circular cylinder: (a) coarse (30 & 60), (b) medium (60 & 120), and (c) fine (120 & 240). The bottom edge is a line of symmetry. Only a portion of each computational domain is shown—the domain extends well beyond the portion shown here.
(c)
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837 CHAPTER 15
longer. A fourth calculation (not shown) at even finer grid resolution shows the same trend—the separation bubbles stretch downstream and the drag coefficient decreases somewhat. Shown in Fig. 15–36 is a contour plot of tangential velocity component (uu) for the medium resolution case. We plot values of uu over a very small range around zero, so that we can clearly see where along the cylinder the flow changes direction. This is thus a clever way to locate the separation point along a cylinder wall. Note that this works only for a circular cylinder because of its unique geometry. A more general way to determine the separation point is to identify the point along the wall where the wall shear stress tw is zero; this technique works for bodies of any shape. From Fig. 15–36, we see that the boundary layer separates at an angle of a " 110° from the front stagnation point, much further downstream than the experimentally obtained value of 82°. In fact, all our CFD results predict boundary layer separation on the rear side rather than the front side of the cylinder. These CFD results are unphysical—such elongated separation bubbles could not remain stable in a real flow situation, the separation point is too far downstream, and the drag coefficient is too low compared to experimental data. Furthermore, repeated grid refinement does not lead to better results as we would hope; on the contrary, the results get worse with grid refinement. Why do these CFD simulations yield such poor agreement with experiment? The answer is twofold: 1. We have forced the CFD solution to be steady, when in fact flow over a circular cylinder at this Reynolds number is not steady. Experiments show that a periodic Kármán vortex street forms behind the cylinder (Tritton, 1977; see also Fig. 4–25 of this text). 2. All three cases in Fig. 15–35 are solved for the upper half-plane only, and symmetry is enforced about the x-axis. In reality, flow over a circular cylinder is highly nonsymmetric; vortices are shed alternately from the top and the bottom of the cylinder, forming the Kármán vortex street. To correct both of these problems, we need to run an unsteady CFD simulation with a full grid (top and bottom)—without imposing the symmetry condition. We run the simulation as an unsteady two-dimensional laminar flow, using the computational domain sketched in Fig. 15–37. The top and bottom (far field) edges are specified as a periodic boundary condition pair so that nonsymmetric oscillations in the wake are not suppressed (flow can cross these boundaries as necessary). The far field boundaries are also very far away (75 to 200 cylinder diameters), so that their effect on the calculations is insignificant. The mesh is very fine near the cylinder to resolve the boundary layer. The grid is also fine in the wake region to resolve the shed vortices as they travel downstream. For this particular simulation, we use a hybrid grid somewhat like that shown in Fig. 15–14. The fluid is air, the cylinder diameter is 1.0 m, and the free-stream air speed is set to 0.00219 m/s. These values produce a Reynolds number of 150 based on cylinder diameter. Note that the Reynolds number is the important parameter in this problem—the choices of D, V, and type of fluid are not critical, so long as they produce the desired Reynolds number (Fig. 15–38).
(a)
(b)
(c)
FIGURE 15–35 Streamlines produced by steady-state CFD calculations of flow over a circular cylinder at Re " 150: (a) coarse grid (30 & 60), (b) medium grid (60 & 120), and (c) fine grid (120 & 240). Note that only the top half of the flow is calculated—the bottom half is displayed as a mirror image of the top. Separation point
y uu < 0
a
u
uu > 0 x
FIGURE 15–36 Contour plot of tangential velocity component uu for flow over a circular cylinder at Re " 150 and for the medium grid resolution case (60 & 120). Values in the range #10#4 ' uu ' 10#4 m/s are plotted, so as to reveal the precise location of boundary layer separation, i.e., where uu changes sign just outside the cylinder wall, as sketched. For this case, the flow separates at a " 110°.
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838 FLUID MECHANICS
V
Far field edge (periodic)
Far field inflow (velocity inlet) Cylinder surface (wall)
Far field outflow (pressure outlet)
y
x
FIGURE 15–37 Computational domain (light blue shaded region) used to simulate unsteady, two-dimensional, laminar flow over a circular cylinder (not to scale). Applied boundary conditions are in parentheses.
Reynolds number is defined as Re =
VD rVD = n m
for flow at free-stream speed V over a circular cylinder of diameter D in a fluid of density r and dynamic viscosity m (kinematic viscosity n).
FIGURE 15–38 In an incompressible CFD simulation of flow around a cylinder, the choice of free-stream speed, cylinder diameter, or even type of fluid is not critical, provided that the desired Reynolds number is achieved.
75D
D
200D
Far field edge (periodic)
As we march in time, small nonuniformities in the flow field amplify, and the flow becomes unsteady and antisymmetric with respect to the xaxis. A Kármán vortex street forms naturally. After sufficient CPU time, the simulated flow settles into a periodic vortex shedding pattern, much like the real flow. A contour plot of vorticity at one instant of time is shown in Fig. 15–39, along with a photograph showing streaklines of the same flow obtained experimentally in a wind tunnel. It is clear from the CFD simulation that the Kármán vortices decay downstream, since the magnitude of vorticity in the vortices decreases with downstream distance. This decay is partly physical (viscous), and partly artificial (numerical dissipation). Nevertheless, physical experiments verify the decay of the Kármán vortices. The decay is not so obvious in the streakline photograph (Fig. 15–39b); this is due to the time-integrating property of streaklines, as was pointed out in Chap. 4. A close-up view of vortices shedding from the cylinder at a particular instant in time is shown in Fig. 15–40, again with a comparison between CFD results and experimental results—this time from experiments in a water channel. A full-color animated version of Fig. 15–40 is provided on the book’s website so that you can watch the dynamic process of vortex shedding. We compare the CFD results to experimental results in Table 15–2. The calculated time-averaged drag coefficient on the cylinder is 1.14. As mentioned previously, experimental values of CD at this Reynolds number vary from about 1.1 to 1.4, so the agreement is within the experimental scatter. Note that the present simulation is two-dimensional, inhibiting any kind of oblique vortex shedding or other three-dimensional nonuniformities. This may be why our calculated drag coefficient is on the lower end of the reported experimental range. The Strouhal number of the Kármán vortex street is defined as Strouhal number:
St "
f sheddingD V
(15–4)
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839 CHAPTER 15 y D
x/D 0
10
30
20
40
50 (a)
60
70
80
90
100
y D
x/D 0
10
20
30
40
50 (b)
60
70
80
90
100
FIGURE 15–39 Laminar flow in the wake of a circular cylinder at Re " 150: (a) an instantaneous snapshot of vorticity contours produced by CFD, and (b) time-integrated streaklines produced by a smoke wire located at x/D " 5. The vorticity contours show that Kármán vortices decay rapidly in the wake, whereas the streaklines retain a “memory” of their history from upstream, making it appear that the vortices continue for a great distance downstream. Photo from Cimbala et al., 1988.
y D
x/D 0
1
2
3
4
5
6
7
8
9
10
11
(a)
y D
x/D 0
1
2
3
4
5
6 (b)
7
8
9
10
11
FIGURE 15–40 Close-up view of vortices shedding from a circular cylinder: (a) instantaneous vorticity contour plot produced by CFD at Re " 150, and (b) dye streaklines produced by dye introduced at the cylinder surface at Re " 140. An animated version of this CFD picture is available on the book’s website. Photo (b) reprinted by permission of Sadatoshi Taneda.
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TA B L E 1 5 – 2 Comparison of CFD results and experimental results for unsteady laminar flow over a circular cylinder at Re " 150* Experiment CFD
CD
St
1.1 to 1.4 1.14
0.18 0.16
* The main cause of the disagreement is most likely due to three-dimensional effects rather than grid resolution or numerical issues.
where fshedding is the shedding frequency of the vortex street. From our CFD simulation, we calculate St " 0.16. The experimentally obtained value of Strouhal number at this Reynolds number is about 0.18 (Williamson, 1989), so again the agreement is reasonable, although the CFD results are a bit low compared to experiment. Perhaps a finer grid would help somewhat, but the major reason for the discrepancy is more likely due to unavoidable threedimensional effects in the experiments, which are not present in these twodimensional simulations. Overall this CFD simulation is a success, as it captures all the major physical phenomena in the flow field. This exercise with “simple” laminar flow over a circular cylinder has demonstrated some of the capabilities of CFD, but has also revealed several aspects of CFD about which one must be cautious. Poor grid resolution can lead to incorrect solutions, particularly with respect to boundary layer separation, but continued refinement of the grid does not lead to more physically correct results if the boundary conditions are not set appropriately (Fig. 15–41). For example, forced numerical flow symmetry is not always wise, even for cases in which the physical geometry is entirely symmetric. Symmetric geometry does not guarantee symmetric flow.
In addition, forced steady flow may yield incorrect results when the flow is inherently unstable and/or oscillatory. Likewise, forced two-dimensionality may yield incorrect results when the flow is inherently three-dimensional. How then can we ensure that a laminar CFD calculation is correct? Only by systematic study of the effects of computational domain size, grid resolution, boundary conditions, flow regime (steady or unsteady, 2-D or 3-D, etc.), along with experimental validation. As with most other areas of engineering, experience is of paramount importance.
15–3
FIGURE 15–41 Poor grid resolution can lead to incorrect CFD results, but a finer grid does not guarantee a more physically correct solution. If the boundary conditions are not specified properly, the results may be unphysical, regardless of how fine the grid.
■
TURBULENT CFD CALCULATIONS
CFD simulations of turbulent flow are much more difficult than those of laminar flow, even for cases in which the flow field is steady in the mean (statisticians refer to this condition as stationary). The reason is that the finer features of the turbulent flow field are always unsteady and threedimensional—random, swirling, vortical structures called turbulent eddies of all orientations arise in a turbulent flow (Fig. 15–42). Some CFD calculations use a technique called direct numerical simulation (DNS), in which an attempt is made to resolve the unsteady motion of all the scales of the turbulent flow. However, the size difference and the time scale difference between the largest and smallest eddies can be several orders of magnitude (L ** h in Fig. 15–42). Furthermore, these differences increase with the Reynolds number (Tennekes and Lumley, 1972), making DNS calculations of turbulent flows even more difficult as the Reynolds number increases. DNS solutions require extremely fine, fully three-dimensional grids, large computers, and an enormous amount of CPU time. With today’s computers, DNS results are not yet feasible for practical high Reynolds number turbulent flows of engineering interest such as flow over a full-scale airplane. This situation is not expected to change for several more decades, even if the fantastic rate of computer improvement continues at today’s pace.
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Thus, we find it necessary to make some simplifying assumptions in order to simulate complex, high Reynolds number, turbulent flow fields. The next level below DNS is large eddy simulation (LES). With this technique, large unsteady features of the turbulent eddies are resolved, while smallscale dissipative turbulent eddies are modeled (Fig. 15–43). The basic assumption is that the smaller turbulent eddies are isotropic; i.e., it is assumed that the small eddies are independent of coordinate system orientation and always behave in a statistically similar and predictable way, regardless of the turbulent flow field. LES requires significantly less computer resources than does DNS, because we eliminate the need to resolve the smallest eddies in the flow field. In spite of this, the computer requirements for practical engineering analysis and design are nevertheless still formidable using today’s technology. Further discussion about DNS and LES is beyond the scope of the present text, but these are areas of much current research. The next lower level of sophistication is to model all the unsteady turbulent eddies with some kind of turbulence model. No attempt is made to resolve the unsteady features of any of the turbulent eddies, not even the largest ones (Fig. 15–44). Instead, mathematical models are employed to take into account the enhanced mixing and diffusion caused by turbulent eddies. For simplicity, we consider only steady (that is, stationary), incompressible flow. When using a turbulence model, the steady Navier–Stokes equation (Eq. 15–2) is replaced by what is called the Reynolds-averaged Navier–Stokes (RANS) equation, shown here for steady (stationary), incompressible, turbulent flow, Steady RANS equation:
→ → → → → 1→ (V ! §)V " # §P$ % n§ 2V % § ! (tij, turbulent) r
h
FIGURE 15–42 All turbulent flows, even those that are steady in the mean (stationary), contain unsteady, three-dimensional turbulent eddies of various sizes. Shown is the average velocity profile and some of the eddies; the smallest turbulent eddies (size h) are orders of magnitude smaller than the largest turbulent eddies (size L). Direct numerical simulation (DNS) is a CFD technique that simulates all relevant turbulent eddies in the flow.
(15–5)
Compared to Eq. 15–2, there is an additional term on the right side of Eq. 15–5 that accounts for the turbulent fluctuations. tij, turbulent is a tensor known as the specific Reynolds stress tensor, so named because it acts in a similar fashion as the viscous stress tensor tij (Chap. 9). In Cartesian coordinates, tij, turbulent is u$2 u$v$ u$w$ tij,turbulent " # £ u$v$ v$2 v$w$ ≥ u$w$ v$w$ w$2
L
(15–6)
where the overbar indicates the time average of the product of two fluctuating velocity components and primes denote fluctuating velocity components. Since the Reynolds stress is symmetric, six additional unknowns are introduced into the problem. These new unknowns are modeled in various ways by turbulence models. A detailed description of turbulence models is beyond the scope of this text; you are referred to Wilcox (1998) or Chen and Jaw (1998) for further discussion. There are many turbulence models in use today, including algebraic, one-equation, two-equation, and Reynolds stress models. Three of the most popular turbulence models are the k-e model, the k-v model, and the q-v model. These so-called two-equation turbulence models add two more transport equations, which must be solved simultaneously with the equations of mass and linear momentum (and also energy if this equation is
FIGURE 15–43 Large eddy simulation (LES) is a simplification of direct numerical simulation in which only the large turbulent eddies are resolved—the small eddies are modeled, significantly reducing computer requirements. Shown is the average velocity profile and the resolved eddies.
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FIGURE 15–44 When a turbulence model is used in a CFD calculation, all the turbulent eddies are modeled, and only Reynolds-averaged flow properties are calculated. Shown is the average velocity profile. There are no resolved turbulent eddies.
Velocity inlet: •V •I •O
D
FIGURE 15–45 A useful rule of thumb for turbulence properties at a pressure inlet or velocity inlet boundary condition is to specify a turbulence intensity of 10 percent and a turbulent length scale of one-half of some characteristic length scale in the problem (! " D/2).
being used). Along with the two additional transport equations that must be solved when using a two-equation turbulence model, two additional boundary conditions must be specified for the turbulence properties at inlets and at outlets. (Note that the properties specified at outlets are not used unless reverse flow is encountered at the outlet.) For example, in the k-e model you may specify both k (turbulent kinetic energy) and e (turbulent dissipation rate). However, appropriate values of these variables are not always known. A more useful option is to specify turbulence intensity I (ratio of characteristic turbulent eddy velocity to free-stream velocity or some other characteristic or average velocity) and turbulent length scale ! (characteristic length scale of the energy-containing turbulent eddies). If detailed turbulence data are not available, a good rule of thumb at inlets is to set I to 10 percent and to set ! to one-half of some characteristic length scale in the flow field (Fig. 15–45). We emphasize that turbulence models are approximations that rely heavily on empirical constants for mathematical closure of the equations. The models are calibrated with the aid of direct numerical simulation and experimental data obtained from simple flow fields like flat plate boundary layers, shear layers, and isotropic decaying turbulence downstream of screens. Unfortunately, no turbulence model is universal, meaning that although the model works well for flows similar to those used for calibration, it is not guaranteed to yield a physically correct solution when applied to general turbulent flow fields, especially those involving flow separation and reattachment and/or large-scale unsteadiness. Turbulent flow CFD solutions are only as good as the appropriateness and validity of the turbulence model used in the calculations.
We emphasize also that this statement remains true regardless of how fine we make the computational grid. When applying CFD to laminar flows, we can often improve the physical accuracy of the simulation by refining the grid. This is not always the case for turbulent flow CFD analyses using turbulence models. While a refined grid produces better numerical accuracy, the physical accuracy of the solution is always limited by the physical accuracy of the turbulence model itself. With these cautions in mind, we now present some practical examples of CFD calculations of turbulent flow fields. In all the turbulent flow examples discussed in this chapter, we employ the k-e turbulence model with wall functions. This model is the default turbulence model in many commercial CFD codes such as FLUENT. In all cases we assume stationary flow; no attempt is made to model unsteady features of the flow, such as vortex shedding in the wake of a bluff body. It is assumed that the turbulence model accounts for all the inherent unsteadiness due to turbulent eddies in the flow field. Note that unsteady (nonstationary) turbulent flows are also solvable with turbulence models, through the use of time-marching schemes (unsteady RANS calculations), but only when the time scale of the unsteadiness is much longer than that of individual turbulent eddies. For example, suppose you are calculating the forces and moments on a blimp during a gust of wind (Fig. 15–46). At the inlet boundary, you would impose the time-varying wind velocity and turbulence levels, and an unsteady turbulent flow solution could then be calculated using turbulence models. The large-
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scale, overall features of the flow (flow separation, forces and moments on the body, etc.) would be unsteady, but the fine-scale features of the turbulent boundary layer, for example, would be modeled by the quasi-steady turbulence model.
FL FD
V(t)
Flow around a Circular Cylinder at Re ! 10,000
As our first example of a turbulent flow CFD solution, we calculate flow over a circular cylinder at Re ! 10,000. For illustration, we use the same two-dimensional computational domain that was used for the laminar cylinder flow calculations, as sketched in Fig. 15–33. As with the laminar flow calculation, only the upper half of the flow field is solved here, due to symmetry along the bottom edge of the computational domain. We use the same three grids used for the laminar flow case as well—coarse, medium, and fine resolution (Fig. 15–34). We point out, however, that grids designed for turbulent flow calculations (especially those employing turbulence models with wall functions) are generally not the same as those designed for laminar flow of the same geometry, particularly near walls. We apply a free-stream flow of air at 25°C and at velocity V ! 7.304 m/s from left to right around this circular cylinder. The Reynolds number of the flow, based on cylinder diameter (D ! 2.0 cm), is approximately 10,000. Experiments at this Reynolds number reveal that the boundary layer is laminar and separates several degrees upstream of the top of the cylinder (at a ! 82°). The wake, however, is turbulent; such a mixture of laminar and turbulent flow is particularly difficult for CFD codes. The measured drag coefficient at this Reynolds number is CD ! 1.15 (Tritton, 1977). CFD solutions are obtained for each of the three grids, assuming stationary (steady in the mean) turbulent flow. We employ the k-e turbulence model with wall functions. The inlet turbulence level is set to 10 percent with a length scale of 0.01 m (half of the cylinder diameter). All three cases converge nicely. Streamlines are plotted in Fig. 15–47 for the three grid resolution cases. In each plot, the image is mirrored about the symmetry line so that even though only the top half of the flow field is solved, the full flow field is visualized. For the coarse resolution case (Fig. 15–47a), the boundary layer separates well past the top of the cylinder, at a " " 140°. Furthermore, the drag coefficient CD is only 0.647, almost a factor of 2 smaller than it should be. Let’s see if a finer mesh improves the agreement with experimental data. For the medium resolution case (Fig. 15–47b), the flow field is significantly different. The boundary layer separates nearer to the top of the cylinder, at a ! 104°, and CD has increased to about 0.742—closer, but still significantly less than the experimental value. We also notice that the recirculating eddies in the cylinder’s wake have grown in length by nearly a factor of 2 compared to those of the coarse grid case. Figure 15–47c shows streamlines for the fine resolution case. The results look very similar to those of the medium resolution case, and the drag coefficient has increased only slightly (CD ! 0.753). The boundary layer separation point for this case is at a ! 102°. Further grid refinement (not shown) does not change the results significantly from those of the fine grid case. In other words, the fine grid appears to be sufficiently resolved, yet the results do not agree with experiment.
FIGURE 15–46 While most CFD calculations with turbulence models are stationary (steady in the mean), it is also possible to calculate unsteady turbulent flow fields using turbulence models. In the case of flow over a body, we may impose unsteady boundary conditions and march in time to predict gross features of the unsteady flow field.
(a)
(b)
(c)
FIGURE 15–47 Streamlines produced by CFD calculations of stationary turbulent flow over a circular cylinder at Re ! 10,000: (a) coarse grid (30 # 60), (b) medium grid (60 # 120), and (c) fine grid (120 # 240). Note that only the top half of the flow is calculated—the bottom half is displayed as a mirror image of the top.
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Why? There are several problems with our calculations: we are modeling a steady flow, even though the actual physical flow is unsteady; we are enforcing symmetry about the x-axis, even though the physical flow is unsymmetric (a Kármán vortex street can be observed in experiments at this Reynolds number); and we are using a turbulence model instead of resolving all the small eddies of the turbulent flow. Another significant source of error in our calculations is that the CFD code is run with turbulence turned on in order to reasonably model the wake region, which is turbulent; however, the boundary layer on the cylinder surface is actually still laminar. The predicted location of the separation point downstream of the top of the cylinder is more in line with turbulent boundary layer separation, which does not occur until much higher values of Reynolds number (after the “drag crisis” at Re greater than 2 & 105). The bottom line is that CFD codes have a hard time in the transitional regime between laminar and turbulent flow, and when there is a mixture of laminar and turbulent flow in the same computational domain. In fact, most commercial CFD codes give the user a choice between laminar and turbulent—there is no “middle ground.” In the present calculations, we model a turbulent boundary layer, even though the physical boundary layer is laminar; it is not surprising, then, that the results of our calculations do not agree well with experiment. If we would have instead specified laminar flow over the entire computational domain, the CFD results would have been even worse (less physical). Is there any way around this problem of poor physical accuracy for the case of mixed laminar and turbulent flow? Perhaps. In some CFD codes you can specify the flow to be laminar or turbulent in different regions of the flow. But even then, the transitional process from laminar to turbulent flow is somewhat abrupt, again not physically correct. Furthermore, you would need to know where the transition takes place in advance—this defeats the purpose of a stand-alone CFD calculation for fluid flow prediction. Advanced wall treatment models are being generated that may some day do a better job in the transitional region. In addition, some new turbulence models are being developed that are better tuned to low Reynolds number turbulence. In summary, we cannot accurately model the mixed laminar/turbulent flow problem of flow over a circular cylinder at Re # 10,000 using standard turbulence models and the steady Reynolds-averaged Navier–Stokes (RANS) equation. It appears that accurate results can be obtained only through use of time-accurate (unsteady RANS), LES, or DNS solutions that are orders of magnitude more computationally demanding.
Flow around a Circular Cylinder at Re " 107
As a final cylinder example, we use CFD to calculate flow over a circular cylinder at Re " 107—well beyond the drag crisis. The cylinder for this case is of 1.0 m diameter, and the fluid is water. The free-stream velocity is 10.05 m/s. At this value of Reynolds number the experimentally measured value of drag coefficient is around 0.7 (Tritton, 1977). The boundary layer is turbulent at the separation point, which occurs at around 120°. Thus we do not have the mixed laminar/turbulent boundary layer problem as in the
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lower Reynolds number example—the boundary layer is turbulent everywhere except near the nose of the cylinder, and we should expect better results from the CFD predictions. We use a two-dimensional half-grid similar to that of the fine resolution case of the previous examples, but the mesh near the cylinder wall is adapted appropriately for this high Reynolds number. As previously, we use the k-e turbulence model with wall functions. The inlet turbulence level is set to 10 percent with a length scale of 0.5 m. Unfortunately, the drag coefficient is calculated to be 0.262—less than half of the experimental value at this Reynolds number. Streamlines are shown in Fig. 15–48. The boundary layer separates a bit too far downstream, at a " 129°. There are several possible reasons for the discrepancy. We are forcing the simulated flow to be steady and symmetric, whereas the actual flow is neither, due to vortex shedding. (Vortices are shed even at high Reynolds numbers.) In addition, the turbulence model and its near wall treatment (wall functions) may not be capturing the proper physics of the flow field. Again we must conclude that accurate results for flow over a circular cylinder can be obtained only through use of a full grid rather than a half grid, and with time-accurate (unsteady RANS), LES, or DNS solutions that are orders of magnitude more computationally demanding.
FIGURE 15–48 Streamlines produced by CFD calculations of stationary turbulent flow over a circular cylinder at Re " 107. Unfortunately, the predicted drag coefficient is still not accurate for this case.
Vane tip Vane hub
Design of the Stator for a Vane-Axial Flow Fan
The next turbulent flow CFD example involves design of the stator for a vane-axial flow fan that is to be used to drive a wind tunnel. The overall fan diameter is D " 1.0 m, and the design point of the fan is at an axial-flow speed of V " 50 m/s. The stator vanes span from radius r " rhub " 0.25 m at the hub to r " rtip " 0.50 m at the tip. The stator vanes are upstream of the rotor blades in this design (Fig. 15–49). A preliminary stator vane shape is chosen that has a trailing edge angle of bst " 63° and a chord length of 20 cm. At any value of radius r, the actual amount of turning depends on the number of stator vanes—we expect that the fewer the number of vanes, the smaller the average angle at which the flow is turned by the stator vanes because of the greater spacing between vanes. It is our goal to determine the minimum number of stator vanes required so that the flow impinging on the leading edges of the rotor blades (located one chord length downstream of the stator vane trailing edges) is turned at an average angle of at least 45°. We also require there to be no significant flow separation from the stator vane surface. As a first approximation, we model the stator vanes at any desired value of r as a two-dimensional cascade of vanes (see Chap. 14). Each vane is separated by blade spacing s at this radius, as defined in Fig. 15–50. We use CFD to predict the maximum allowable value of s, from which we estimate the minimum number of stator vanes that meet the given requirements of the design. Since the flow through the two-dimensional cascade of stator vanes is infinitely periodic in the y-direction, we need to model only one flow passage through the vanes, specifying two pairs of periodic boundary conditions on the top and bottom edges of the computational domain (Fig. 15–51). We run six cases, each with a different value of blade spacing. We choose s " 10, 20, 30, 40, 50, and 60 cm, and generate a structured grid for
Hub and motor vr
bst D r
V
Stator
v
Rotor
FIGURE 15–49 Schematic diagram of the vane-axial flow fan being designed. The stator precedes the rotor, and the flow through the stator vanes is to be modeled with CFD.
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FIGURE 15–50 Definition of blade spacing s: (a) frontal view of the stator, and (b) the stator modeled as a twodimensional cascade in edge view. Twelve radial stator vanes are shown in the frontal view, but the actual number of vanes is to be determined. Three stator vanes are shown in the cascade, but the actual cascade consists of an infinite number of vanes, each displaced by blade spacing s, which increases with radius r. The two-dimensional cascade is an approximation of the threedimensional flow at one value of radius r and blade spacing s. Chord length c is defined as the horizontal length of the stator vane.
Translationally periodic 1 y x V
s
Translationally periodic 2
Pressure surface Suction surface Pressure outlet
Velocity inlet
FIGURE 15–51 Computational domain (light blue shaded region) defined by one flow passage through two stator vanes. The top wall of the passage is the pressure surface, and the bottom wall is the suction surface. Two translationally periodic pairs are defined: periodic 1 upstream and periodic 2 downstream.
s at r = rtip s
s
y
r x
D s rhub
c
rtip
(a)
(b)
each of these values of blade spacing. The grid for the case with s " 10 cm is shown in Fig. 15–52; the other grids are similar, but more intervals are specified in the y-direction as s increases. Notice how we have made the grid spacing fine near the pressure and suction surfaces so that the boundary layer on these surfaces can be better resolved. We specify V " 50 m/s at the velocity inlet, zero gage pressure at the pressure outlet, and a smooth wall boundary condition with no slip at both the pressure and suction surfaces. Since we are modeling the flow with a turbulence model (k-e with wall functions), we must also specify turbulence properties at the velocity inlet. For these simulations we specify a turbulence intensity of 10 percent and a turbulence length scale of 0.01 m (1.0 cm). We run the CFD calculations long enough to converge as far as possible for all six cases, and we plot streamlines in Fig. 15–53 for six blade spacings: s " 10, 20, 30, 40, 50, and 60 cm. Although we solve for flow through only one flow passage, we draw several duplicate flow passages, stacked one on top of the other, in order to visualize the flow field as a periodic cascade. The streamlines for the first three cases look very similar at first glance, but closer inspection reveals that the average angle of flow downstream of the trailing edge of the stator vane decreases with s. (We define flow angle b relative to horizontal as sketched in Fig. 15–53a.) Also, the gap (white space) between the wall and the closest streamline to the suction surface increases in size as s increases, indicating that the flow speed in that region decreases. In fact, it turns out that the boundary layer on the suction surface of the stator vane must resist an ever-increasingly adverse pressure gradient (decelerating flow speed and positive pressure gradient) as blade spacing is increased. At large enough s, the boundary layer on the suction surface cannot withstand the severely adverse pressure gradient and separates off the wall. For s " 40, 50, and 60 cm (Fig. 15–53d through f ), flow separation off the suction surface is clearly seen in these streamline plots. Furthermore, the severity of the flow separation increases with s. This is not
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FIGURE 15–52 Structured grid for the twodimensional stator vane cascade at blade spacing s " 10 cm. The outflow region in the wake of the vanes is intentionally longer than that at the inlet to avoid backflow at the pressure outlet in case of flow separation on the suction surface of the stator vane. The outlet is one chord length downstream of the stator vane trailing edges; the outlet is also the location of the leading edges of the rotor blades (not shown).
b
(a)
(b)
(c)
(d)
(e)
(f)
FIGURE 15–53 Streamlines produced by CFD calculations of stationary turbulent flow through a stator vane flow passage: (a) blade spacing s " 10, (b) 20, (c) 30, (d) 40, (e) 50, and ( f ) 60 cm. The CFD calculations are performed using the k-e turbulence model with wall functions. Flow angle b is defined in image (a) as the average angle of flow, relative to horizontal, just downstream of the trailing edge of the stator vane.
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TA B L E 1 5 – 3 Variation of average outlet flow angle bavg, average outlet flow speed Vavg, and predicted drag force per unit depth FD /b as functions of blade spacing s* s, cm 10 20 30 40 50 60
bavg, degrees
Vavg, m/s
FD /b, N/m
60.8 56.1 49.7 43.2 37.2 32.3
103 89.6 77.4 68.6 62.7 59.1
554 722 694 612 538 489
* All calculated values are reported to three significant digits. The CFD calculations are performed using the k-e turbulence model with wall functions.
unexpected if we imagine the limit as s → +. In that case, the stator vane is isolated from its neighbors, and we surely expect massive flow separation since the vane has such a high degree of camber. We list average outlet flow angle bavg, average outlet flow speed Vavg, and predicted drag force on a stator vane per unit depth FD /b in Table 15–3 as functions of blade spacing s. (Depth b is into the page of Fig. 15–53 and is assumed to be 1 m in two-dimensional calculations such as these.) While bavg and Vavg decrease continuously with s, FD /b first rises to a maximum for the s " 20 cm case, and then decreases from there on. You may recall from the previously stated design criteria that the average outlet flow angle must be greater than 45°, and there must be no significant flow separation. From our CFD results, it appears that both of these criteria break down somewhere between s " 30 and 40 cm. We obtain a better picture of flow separation by plotting vorticity contours (Fig. 15–54). In these grayscale contour plots, black represents large negative vorticity (clockwise rotation), white represents large positive vorticity (counterclockwise rotation), and middle gray is zero vorticity. If the boundary layer remains attached, we expect the vorticity to be concentrated within thin boundary layers along the
FIGURE 15–54 Vorticity contour plots produced by CFD calculations of stationary turbulent flow through a stator vane flow passage: blade spacing (a) s " 30 cm and (b) s " 40 cm. The flow field is largely irrotational (zero vorticity) except in the thin boundary layer along the walls and in the wake region. However, when the boundary layer separates, as in case (b), the vorticity spreads throughout the separated flow region.
(a)
(b)
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stator vane surfaces, as is the case in Fig. 15–54a for s " 30 cm. If the boundary layer separates, however, the vorticity suddenly spreads out away from the suction surface, as seen in Fig. 15–54b for s " 40 cm. These results verify that significant flow separation occurs somewhere between s " 30 and 40 cm. As a side note, notice how the vorticity is concentrated not only in the boundary layer, but also in the wake for both cases shown in Fig. 15–54. Finally, we compare velocity vector plots in Fig. 15–55 for three cases: s " 20, 40, and 60 cm. We generate several equally spaced parallel lines in the computational domain; each line is tilted at 45° from horizontal. Velocity vectors are then plotted along each of these parallel lines. When s " 20 cm (Fig. 15–55a), the boundary layer remains attached on both the suction and pressure surfaces of the stator vane all the way to its trailing edge. When s " 40 cm (Fig. 15–55b), flow separation and reverse flow along the suction surface appears. When s " 60 cm (Fig. 15–55c), the separation bubble and the reverse flow region have grown – this is a “dead” flow region, in which the air speeds are very small. In all cases, the flow on the pressure surface of the stator vane remains attached. How many vanes (N) does a blade spacing of s " 30 cm represent? We can easily calculate N by noting that at the vane tip (r " rtip " D/2 " 50 cm), where s is largest, the total available circumference (C) is Available circumference:
C " 2prtip " pD
(15–7)
The number of vanes that can be placed within this circumference with a blade spacing of s " 30 cm is thus Maximum number of vanes:
N"
C pD p(100 cm) " " " 10.5 s s 30 cm
(a)
(b)
(15–8)
Obviously we can have only an integer value of N, so we conclude from our preliminary analysis that we should have at least 10 or 11 stator vanes. How good is our approximation of the stator as a two-dimensional cascade of vanes? To answer this question, we perform a full three-dimensional CFD analysis of the stator. Again we take advantage of the periodicity by modeling only one flow passage—a three-dimensional passage between two radial stator vanes (Fig. 15–56). We choose N " 10 stator vanes by specifying an angle of periodicity of 360/10 " 36°. From Eq. 15–8, this represents s " 31.4 at the vane tips and s " 15.7 at the hub, for an average value of savg " 23.6. We generate a hexagonal structured grid in a computational domain bounded by a velocity inlet, an outflow outlet, a section of cylindrical wall at the hub and another at the tip, the pressure surface of the vane, the suction surface of the vane, and two pairs of periodic boundary conditions. In this three-dimensional case, the periodic boundaries are rotationally periodic instead of translationally periodic. Note that we use an outflow boundary condition rather than a pressure outlet boundary condition, because we expect the swirling motion to produce a radial pressure distribution on the outlet face. The grid is finer near the walls than elsewhere (as usual), to better resolve the boundary layer. The incoming velocity, turbulence level, turbulence model, etc., are all the same as those used for the two-dimensional approximation. The total number of computational cells is almost 800,000.
(c)
FIGURE 15–55 Velocity vectors produced by CFD calculations of stationary turbulent flow through a stator vane flow passage: blade spacing (a) s " 20 cm, (b) 40 cm, and (c) 60 cm.
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850 FLUID MECHANICS Pressure surface
FIGURE 15–56 Three-dimensional computational domain defined by one flow passage through two stator vanes for N " 10 (angle between vanes " 36°). The computational domain volume is defined between the pressure and suction surfaces of the stator vanes, between the inner and outer cylinder walls, and from the inlet to the outlet. Two pairs of rotationally periodic boundary conditions are defined as shown.
FIGURE 15–57 Pressure contour plot produced by three-dimensional CFD calculations of stationary turbulent flow through a stator vane flow passage. Pressure is shown in N/m2 on the vane surfaces and the inner cylinder wall (the hub). Outlines of the inlet and outlet are also shown for clarity. Although only one flow passage is modeled in the CFD calculations, we duplicate the image circumferentially around the x-axis nine times to visualize the entire stator flow field. In this grayscale image, high pressures (as on the pressure surfaces of the vanes) are light, while low pressures (as on the suction surfaces of the vanes, especially near the hub) are dark.
Outflow outlet
V Outer cylinder wall
Inner cylinder wall
Suction surface
Velocity inlet Rotationally periodic 1
Rotationally periodic 2
y z
x
3.79e % 03 3.38e % 03 2.98e % 03 2.57e % 03 2.17e % 03 1.77e % 03 1.36e % 03 9.57e % 02 5.53e % 02 1.49e % 02 –2.55e % 02 –6.59e % 02 –1.06e % 03 –1.47e % 03 –1.87e % 03 –2.28e % 03 –2.68e % 03 –3.08e % 03 –3.49e % 03 –3.89e % 03 –4.30e % 03 –4.70e % 03 –5.10e % 03 –5.51e % 03 –5.91e % 03 –6.32e % 03
Pressure surface
V Inlet Suction surface
Outlet
y z
x
Pressure contours on the stator vane surfaces and on the inner cylindrical wall are plotted in Fig. 15–57. This view is from the same angle as that of Fig. 15–56, but we have zoomed out and duplicated the computational domain nine times circumferentially about the axis of rotation (the x-axis) for a total of 10 flow passages to aid in visualization of the flow field. You can see that the pressure is higher (lighter shade of gray) on the pressure surface than on the suction surface. You can also see an overall drop in pressure along the hub surface from upstream to downstream of the stator. The change in average pressure from the inlet to the outlet is calculated to be 3.29 kPa. To compare our three-dimensional results directly with the two-dimensional approximation, we run one additional two-dimensional case at the average blade spacing, s " savg " 23.6 cm. A comparison between the two- and three-
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851 CHAPTER 15
dimensional cases is shown in Table 15–4. From the three-dimensional calculation, the net axial force on one stator vane is FD " 183 N. We compare this to the two-dimensional value by converting to force per unit depth (force per unit span of the stator vane). Since the stator vane spans 0.25 m, FD /b " (183 N)/(0.25 m) " 732 N/m. The corresponding two-dimensional value from Table 15–4 is FD /b " 724 N/m, so the agreement is very good (! 1 percent difference). The average speed at the outlet of the three-dimensional domain is Vavg " 84.7 m/s, almost identical to the two-dimensional value of 84.8 m/s in Table 15–4. The two-dimensional approximation differs by less than 1 percent. Finally, the average outlet flow angle bavg obtained from our full three-dimensional calculation is 53.3°, which easily meets the design criterion of 45°. We compare this to the two-dimensional approximation of 53.9° in Table 15–4; the agreement is again around 1 percent. Contours of tangential velocity component at the outlet of the computational domain are plotted in Fig. 15–58. We see that the tangential velocity distribution is not uniform; it decreases as we move radially outward from hub to tip as we should expect, since blade spacing s increases from hub to tip. We also find (not shown here) that the outflow pressure increases radially from hub to tip. This also agrees with our intuition, since we know that a radial pressure gradient is required to sustain a tangential flow—the pressure rise with increasing radius provides the centripetal acceleration necessary to turn the flow about the x-axis. Another comparison can be made between the three-dimensional and twodimensional calculations by plotting vorticity contours in a slice through the computational domain within the flow passage between vanes. Two such
9.00e % 01 8.70e % 01 8.40e % 01 8.10e % 01 7.80e % 01 7.50e % 01 7.20e % 01 6.90e % 01 6.60e % 01 6.30e % 01 6.00e % 01 5.70e % 01 5.40e % 01 5.10e % 01 4.80e % 01 4.50e % 01 4.20e % 01 3.90e % 01 3.60e % 01 3.30e % 01 3.00e % 01 2.70e % 01 2.40e % 01 2.10e % 01 1.80e % 01 1.50e % 01 1.20e % 01 9.00e % 00 6.00e % 00 3.00e % 00 0.00e % 00
Pressure surface
V Inlet
Suction surface
Outlet y z
x
TA B L E 1 5 – 4 CFD results for flow through a stator vane flow passage: the twodimensional cascade approximation at the average blade spacing, (s " savg " 23.6 cm) is compared to the full three-dimensional calculation* 2-D, s " 23.6 cm Full 3-D bavg Vavg, m/s FD /b, N/m
53.9° 84.8 724
53.3° 84.7 732
* Values are shown to three significant digits.
FIGURE 15–58 Grayscale tangential velocity contour plot produced by three-dimensional CFD calculations of stationary turbulent flow through a stator vane flow passage. The tangential velocity component is shown in m/s at the outlet of the computational domain (and also on the vane surfaces, where the velocity is zero). An outline of the inlet to the computational domain is also shown for clarity. Although only one flow passage is modeled, we duplicate the image circumferentially around the x-axis nine times to visualize the entire stator flow field. In this grayscale image, the tangential velocity ranges from 0 (black) to 90 m/s (white).
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852 FLUID MECHANICS
slices are created—a slice close to the hub and a slice close to the tip, and vorticity contours are plotted in Fig. 15–59. In both slices, the vorticity is confined to the thin boundary layer and wake. There is no flow separation near the hub, but we see that near the tip, the flow has just begun to separate on the suction surface near the trailing edge of the stator vane. Notice that the air leaves the trailing edge of the vane at a steeper angle at the hub than at the tip. This also agrees with our two-dimensional approximation (and our intuition), since blade spacing s at the hub (15.7 cm) is smaller than s at the tip (31.4 cm). In conclusion, the approximation of this three-dimensional stator as a two-dimensional cascade of stator vanes turns out to be quite good overall, particularly for preliminary analysis. The discrepancy between the two- and three-dimensional calculations for gross flow features, such as force on the vane, outlet flow angle, etc., is around 1 percent or less for all reported quantities. It is therefore no wonder that the two-dimensional cascade approach is such a popular approximation in turbomachinery design. The more detailed three-dimensional analysis gives us confidence that a stator
FIGURE 15–59 Vorticity contour plots produced by three-dimensional stationary turbulent CFD calculations of flow through a stator vane flow passage: (a) a slice near the hub or root of the vanes and (b) a slice near the tip of the vanes. Contours of z-vorticity are plotted, since the faces are nearly perpendicular to the z-axis. In these grayscale images, very dark regions (as in the upper half of the wake and in the flow separation zone) represent negative (clockwise) z-vorticity, while very light regions (as in the lower half of the wake) represent positive (counterclockwise) z-vorticity. Near the hub, there is no sign of flow separation, but near the tip, there is some indication of flow separation near the trailing edge of the suction side of the vane. Also shown are arrows indicating how the periodic boundary condition works. Flow leaving the bottom of the periodic boundary enters at the same speed and direction into the top of the periodic boundary. Outflow angle b is larger near the hub than near the tip of the stator vanes, because blade spacing s is smaller at the hub than at the tip, and also because of the mild flow separation near the tip.
Inlet
V Outlet Pressure surface b y z
Suction surface
x
(a) Inlet
V
Flow separation
Pressure surface
Outlet
b
y z
x
Suction surface
(b)
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853 CHAPTER 15
with 10 vanes is sufficient to meet the imposed design criteria for this axialflow fan. However, our three-dimensional calculations have revealed a small separated region near the tip of the stator vane. It may be wise to apply some twist to the stator vanes (reduce the pitch angle or angle of attack toward the tip) in order to avoid this separation. (Twist is discussed in more detail in Chap. 14.) Alternatively, we can increase the number of stator vanes to 11 or 12 to hopefully eliminate flow separation at the vane tips. As a final comment on this example flow field, all the calculations were performed in a fixed coordinate system. Modern CFD codes contain options for modeling zones in the flow field with rotating coordinate systems so that similar analyses can be performed on rotor blades as well as on stator vanes.
15–4
■
q⋅ wall computed
Fluid
Twall specified
Solid (a)
CFD WITH HEAT TRANSFER
By coupling the differential form of the energy equation with the equations of fluid motion, we can use a computational fluid dynamics code to calculate properties associated with heat transfer (e.g., temperature distributions or rate of heat transfer from a solid surface to a fluid). Since the energy equation is a scalar equation, only one extra transport equation (typically for either temperature or enthalpy) is required, and the computational expense (CPU time and RAM requirements) is not increased significantly. Heat transfer capability is built into most commercially available CFD codes, since many practical problems in engineering involve both fluid flow and heat transfer. As mentioned previously, additional boundary conditions related to heat transfer need to be specified. At solid wall boundaries, we . may specify either wall temperature Twall (K) or the wall heat flux qwall 2 (W/m ), defined as the rate of heat transfer per unit area from the wall to the fluid (but not both at the same time, as illustrated in Fig. 15–60). When we model a zone in a computational domain as a solid body that involves the generation of thermal energy via electric heating (as in electronic components) or chemical or nuclear reactions (as in nuclear fuel rods), we may . instead specify the heat generation rate per unit volume g (W/m3) within the solid since the ratio of the total heat generation rate to the exposed surface . area must equal the average wall heat flux. In that case, neither Twall nor qwall are specified; both converge to values that match the specified heat generation rate. In addition, the temperature distribution inside the solid object itself can then be calculated. Other boundary conditions (such as those associated with radiation heat transfer) may also be applied in CFD codes. In this section we do not go into details about the equations of motion or the numerical techniques used to solve them. Rather, we show some basic examples that illustrate the capability of CFD to calculate practical flows of engineering interest that involve heat transfer.
Temperature Rise through a Cross-Flow Heat Exchanger
Consider flow of cool air through a series of hot tubes as sketched in Fig. 15–61. In heat exchanger terminology, this geometrical configuration is called a cross-flow heat exchanger. If the airflow were to enter horizontally (a " 0) at all times, we could cut the computational domain in half and
q⋅ wall specified
Fluid
Twall computed
Solid (b)
FIGURE 15–60 At a wall boundary, we may specify either (a) the wall temperature or (b) the wall heat flux, but not both, as this would be mathematically overspecified.
D
Translationally periodic
Computational domain
In a
Out
3D
3D
Translationally periodic
FIGURE 15–61 The computational domain (light blue shaded region) used to model turbulent flow through a cross-flow heat exchanger. Flow enters from the left at angle a from the horizontal.
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854 FLUID MECHANICS
FIGURE 15–62 Close-up view of the structured grid near one of the cross-flow heat exchanger tubes. The grid is fine near the tube walls so that the wall boundary layer can be better resolved.
FIGURE 15–63 Temperature contour plots produced by CFD calculations of stationary turbulent flow through a cross-flow heat exchanger at a " 0° with smooth tubes. Grayscale contours range from 300 K (darkest) to 315 K or higher (lightest). The average air temperature at the outlet increases by 5.51 K compared to the inlet air temperature. Note that although the calculations are performed in the computational domain of Fig. 15–61, the image is duplicated here three times for purposes of illustration.
apply symmetry boundary conditions on the top and bottom edges of the domain (see Fig. 15–23). In the case under consideration, however, we allow the airflow to enter the computational domain at some angle (a , 0). Thus, we impose translationally periodic boundary conditions on the top and bottom edges of the domain as sketched in Fig. 15–61. We set the inlet air temperature to 300 K and the surface temperature of each tube to 500 K. The diameter of the tubes and the speed of the air are chosen such that the Reynolds number is approximately 1 & 105 based on tube diameter. The tube surfaces are assumed to be hydrodynamically smooth (zero roughness) in this first set of calculations. The hot tubes are staggered as sketched in Fig. 15–61 and are spaced three diameters apart both horizontally and vertically. We assume two-dimensional stationary turbulent flow without gravity effects and set the turbulence intensity of the inlet air to 10 percent. We run two cases for comparison: a " 0 and 10°. Our goal is to see whether the heat transfer to the air is enhanced or inhibited by a nonzero value of a. Which case do you think will provide greater heat transfer? We generate a two-dimensional, multiblock, structured grid with very fine resolution near the tube walls as shown in Fig. 15–62, and we run the CFD code to convergence for both cases. Temperature contours are shown for the a " 0° case in Fig. 15–63, and for the a " 10° case in Fig. 15–64. The average rise of air temperature leaving the outlet of the control volume for the case with a " 0° is 5.51 K, while that for a " 10° is 5.65 K. Thus we conclude that the off-axis inlet flow leads to more effective heating of the air, although the improvement is only about 2.5 percent. We compute a third case (not shown) in which a " 0° but the turbulence intensity of the incoming air is increased to 25 percent. This leads to improved mixing, and the average air temperature rise from inlet to outlet increases by about 6.5 percent to 5.87 K. Finally, we study the effect of rough tube surfaces. We model the tube walls as rough surfaces with a characteristic roughness height of 0.01 m (1 percent of cylinder diameter). Note that we had to coarsen the grid somewhat near each tube so that the distance from the center of the closest computational cell to the wall is greater than the roughness height; otherwise the
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855 CHAPTER 15
FIGURE 15–64 Temperature contour plots produced by CFD calculations of stationary turbulent flow through a cross-flow heat exchanger at a " 10° with smooth tubes. Grayscale contours range from 300 K (darkest) to 315 K or higher (lightest). The average air temperature at the outlet increases by 5.65 K compared to the inlet air temperature. Thus, off-axis inlet flow (a " 10°) yields a )T that is 2.5 percent higher than that for the on-axis inlet flow (a " 0°). FIGURE 15–65 Temperature contour plots produced by CFD calculations of stationary turbulent flow through a cross-flow heat exchanger at a " 0° with rough tubes (average wall roughness equal to 1 percent of tube diameter; wall functions utilized in the CFD calculations). Grayscale contours range from 300 K (darkest) to 315 K or higher (lightest). The average air temperature at the outlet increases by 14.48 K compared to the inlet air temperature. Thus, even this small amount of surface roughness yields a )T that is 163 percent higher than that for the case with smooth tubes.
roughness model in the CFD code is unphysical. The flow inlet angle is set to a " 0° for this case, and flow conditions are identical to those of Fig. 15–63. Temperature contours are plotted in Fig. 15–65. Pure white regions in the contour plot represent locations where the air temperature is greater than 315 K. The average air temperature rise from inlet to outlet is 14.48 K, a 163 percent increase over the smooth wall case at a " 0°. Thus we see that wall roughness is a critical parameter in turbulent flows. This example provides some insight as to why the tubes in heat exchangers are often purposely roughened.
Cooling of an Array of Integrated Circuit Chips
In electronics equipment, instrumentation, and computers, electronic components, such as integrated circuits (ICs or “chips”), resistors, transistors, diodes, and capacitors, are soldered onto printed circuit boards (PCBs).
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856 FLUID MECHANICS PCB IC
Cooling air at V " 2.60 m/s and T∞ " 30°C
FIGURE 15–66 Four printed circuit boards (PCBs) stacked in rows, with air blown in between each PCB to provide cooling.
The PCBs are often stacked in rows as sketched in Fig. 15–66. Because many of these electronic components must dissipate heat, cooling air is often blown through the air gap between each pair of PCBs to keep the components from getting too hot. Consider the design of a PCB for an outer space application. Several identical PCBs are to be stacked as in Fig. 15–66. Each PCB is 10 cm high and 30 cm long, and the spacing between boards is 2.0 cm. Cooling air enters the gap between the PCBs at a speed of 2.60 m/s and a temperature of 30°C. The electrical engineers must fit eight identical ICs on a 10 cm & 15 cm portion of each board. Each of the ICs dissipates 6.24 W of heat: 5.40 W from its top surface and 0.84 W from its sides. (There is assumed to be no heat transfer from the bottom of the chip to the PCB.) The rest of the components on the board have negligible heat transfer compared to that from the eight ICs. To ensure adequate performance, the average temperature on the chip surface should not exceed 150°C, and the maximum temperature anywhere on the surface of the chip should not exceed 180°C. Each chip is 2.5 cm wide, 4.5 cm long, and 0.50 cm thick. The electrical engineers come up with two possible configurations of the eight chips on the PCB as sketched in Fig. 15–67: in the long configuration, the chips are aligned with their long dimension parallel to the flow, and in the short configuration, the chips are aligned with their short dimension parallel to the flow. The chips are staggered in both cases to enhance cooling. We are to determine which arrangement leads to the lower maximum surface temperature on the chips, and whether the electrical engineers will meet the surface temperature requirements. For each configuration, we define a three-dimensional computational domain consisting of a single flow passage through the air gap between two PCBs (Fig. 15–68). We generate a structured hexagonal grid with 267,520 cells for each configuration. The Reynolds number based on the 2.0-cm gap between boards is about 3600. If this were a simple two-dimensional channel flow, this Reynolds number would be barely high enough to establish turbulent flow. However, since the surfaces leading up to the velocity inlet are very rough, the flow is most likely turbulent. We note that low Reynolds number turbulent flows are challenging for most turbulence models, since the models are calibrated at high Reynolds numbers. Nevertheless, we
Long configuration
FIGURE 15–67 Two possible configurations of the eight ICs on the PCB: long configuration and short configuration. Without peeking ahead, which configuration do you think will offer the best cooling to the chips?
Short configuration
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857 CHAPTER 15 Short configuration
Long configuration
Velocity inlet
8 3 2
5
7
4
x
5
z
7
3 4 Pressure outlet
Adiabatic walls
6
2
6
1 y
8
Velocity inlet
y
1
x
Pressure outlet Adiabatic walls
z
assume stationary turbulent flow and employ the k-e turbulence model with wall functions. While the absolute accuracy of these calculations may be suspect because of the low Reynolds number, comparisons between the long and short configurations should be reasonable. We ignore buoyancy effects in the calculations since this is a space application. The inlet is specified as a velocity inlet (air) with V " 2.60 m/s and T+ " 30°C; we set the inlet turbulence intensity to 20 percent and the turbulent length scale to 1.0 mm. The outlet is a pressure outlet at zero gage pressure. The PCB is modeled as a smooth adiabatic wall (zero heat transfer from the wall to the air). The top and sides of the computational domain are also approximated as smooth adiabatic walls. Based on the given chip dimensions, the surface area of the top of a chip is 4.5 cm & 2.5 cm " 11.25 cm2. The total surface area of the four sides of the chip is 7.0 cm2. From the given heat transfer rates, we calculate the rate of heat transfer per unit area from the top surface of each chip, # qtop "
5.4 W " 0.48 W/cm2 11.25 cm2
So, we model the top surface of each chip as a smooth wall with a surface heat flux of 4800 W/m2 from the wall to the air. Similarly, the rate of heat transfer per unit area from the sides of each chip is 0.84 W # qsides " " 0.12 W/cm2 7.0 cm2
Since the sides of the chip have electrical leads, we model each side surface of each chip as a rough wall with an equivalent roughness height of 0.50 mm and a surface heat flux of 1200 W/m2 from the wall to the air. The CFD code FLUENT is run for each case to convergence. Results are summarized in Table 15–5, and temperature contours are plotted in Figs. 15–69 and 15–70. The average temperature on the top surfaces of the chips is about the same for either configuration (144.4°C for the long case and 144.7°C for the short case) and is below the recommended limit of 150°C.
FIGURE 15–68 Computational domains for the chip cooling example. Air flowing through the gap between two PCBs is modeled. Two separate grids are generated, one for the long configuration and one for the short configuration. Chips 1 through 8 are labeled for reference. The surfaces of these chips transfer heat to the air; all other walls are adiabatic.
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858 FLUID MECHANICS
TA B L E 1 5 – 5 Comparison of CFD results for the chip cooling example, long and short configurations Tmax, top surfaces of chips Tavg, top surfaces of chips Tmax, side surfaces of chips Tavg, side surfaces of chips Average )T, inlet to outlet Average )P, inlet to outlet
FIGURE 15–69 CFD results for the chip cooling example, long configuration: grayscale temperature contours as viewed from directly above the chip surfaces, with T values in K on the legend. The location of maximum surface temperature is indicated, it occurs near the end of chip 7. Light regions near the leading edges of chips 1, 2, and 3 are also seen, indicating high surface temperatures at those locations.
4.62e + 02 4.56e + 02 4.50e + 02 4.44e + 02 4.38e + 02 4.32e + 02 4.26e + 02 4.20e + 02 4.14e + 02 4.08e + 02 4.02e + 02 3.96e + 02 3.90e + 02 3.84e + 02 3.78e + 02 3.72e + 02 3.66e + 02 3.60e + 02 3.54e + 02 3.48e + 02 3.42e + 02 3.36e + 02 3.30e + 02 3.24e + 02 3.18e + 02 3.12e + 02 3.06e + 02 3.00e + 02
Long
Short
187.5°C 144.5°C 154.0°C 84.2°C 7.83°C #5.14 Pa
182.1°C 144.7°C 170.6°C 91.4°C 7.83°C #5.58 Pa
Long configuration Tmax = 460.7 K 8
3 5 2
7 4
1
6
Airflow direction
y z
x
There is more of a difference in average temperature on the side surfaces of the chips, however (84.2°C for the long case and 91.4(C for the short case), although these values are well below the limit. Of greatest concern are the maximum temperatures. For the long configuration, Tmax " 187.5°C and occurs on the top surface of chip 7 (the middle chip of the last row). For the short configuration, Tmax " 182.1°C and occurs close to midboard on the top surfaces of chips 7 and 8 (the two chips in the last row). For both configurations these values exceed the recommended limit of 180°C, although not by much. The short configuration does a better job at cooling the top surfaces of the chips, but at the expense of a slightly larger pressure drop and poorer cooling along the side surfaces of the chips. Notice from Table 15–5 that the average change in air temperature from inlet to outlet is identical for both configurations (7.83°C). This should not be
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859 CHAPTER 15 4.62e + 02 4.56e + 02 4.50e + 02 4.44e + 02 4.38e + 02 4.32e + 02 4.26e + 02 4.20e + 02 4.14e + 02 4.08e + 02 4.02e + 02 3.96e + 02 3.90e + 02 3.84e + 02 3.78e + 02 3.72e + 02 3.66e + 02 3.60e + 02 3.54e + 02 3.48e + 02 3.42e + 02 3.36e + 02 3.30e + 02 3.24e + 02 3.18e + 02 3.12e + 02 3.06e + 02 3.00e + 02
Short configuration Tmax = 455.3 K
2 3 1
8
5 6 4
7
Airflow direction
y z
x
surprising, because the total rate of heat transferred from the chips to the air is the same regardless of chip configuration. In fact, in a CFD analysis it is wise to check values like this—if average )T were not the same between the two configurations, we would suspect some kind of error in our calculations. There are many other interesting features of these flow fields that we can point out. For either configuration, the average surface temperature on the downstream chips is greater than that on the upstream chips. This makes sense physically, since the first chips receive the coolest air, while those downstream are cooled by air that has already been warmed up somewhat. We notice that the front chips (1, 2, and 3 in the long configuration and 1 and 2 in the short configuration) have regions of high temperature just downstream of their leading edges. A close-up view of the temperature distribution on one of these chips is shown in Fig. 15–71a. Why is the temperature so high there? It turns out that the flow separates off the sharp corner at the front of the chip and forms a recirculating eddy called a separation bubble on the top of the chip (Fig. 15–71b). The air speed is slow in that region, especially along the reattachment line where the flow reattaches to the surface. The slow air speed leads to a local “hot spot” in that region of the chip surface since convective cooling is minimal there. Finally, we notice in Fig. 15–71a that downstream of the separation bubble, T increases down the chip surface. There are two reasons for this: (1) the air warms up as it travels down the chip, and (2) the boundary layer on the chip surface grows downstream. The larger the boundary layer thickness, the lower the air speed near the surface, and thus the lower the amount of convective cooling at the surface.
FIGURE 15–70 CFD results for the chip cooling example, short configuration: grayscale temperature contours as viewed from directly above the chip surfaces, with T values in K on the legend. The same temperature scale is used here as in Fig. 15–69. The locations of maximum surface temperature are indicated; they occur near the end of chips 7 and 8 near the center of the PCB. Light regions near the leading edges of chips 1 and 2 are also seen, indicating high surface temperatures at those locations.
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860 FLUID MECHANICS Chip 2, long configuration Cooling air
Region of high T (a)
Approximate location of reattachment line
In summary, our CFD calculations have predicted that the short configuration leads to a lower value of maximum temperature on the chip surfaces and appears at first glance to be the preferred configuration for heat transfer. However, the short configuration demands a higher pressure drop at the same volume flow rate (Table 15–5). For a given cooling fan, this additional pressure drop would shift the operating point of the fan to a lower volume flow rate (Chap. 14), decreasing the cooling effect. It is not known whether this shift would be enough to favor the long configuration—more information about the fan and more analysis would be required. The bottom line in either case is that there is not sufficient cooling to keep the chip surface temperature below 180°C everywhere on every chip. To rectify the situation, we recommend that the designers spread the eight hot chips over the entire PCB rather than in the limited 10 cm & 15 cm area of the board. The increased space between chips should result in sufficient cooling for the given flow rate. Another option is to install a more powerful fan that would increase the speed of the inlet air.
Separation bubble
15–5 (b)
FIGURE 15–71 (a) Close-up top view of grayscale temperature contours on the surface of chip 2 of the long configuration. The region of high temperature is outlined. Temperature contour levels are the same as in Fig. 15–69. (b) An even closer view (an edge view) of streamlines outlining the separation bubble in that region. The approximate location of the reattachment line on the chip surface is also shown.
■
COMPRESSIBLE FLOW CFD CALCULATIONS
All the examples discussed in this chapter so far have been for incompressible flow (r " constant). When the flow is compressible, density is no longer a constant, but becomes an additional variable in the equation set. We limit our discussion here to ideal gases. When we apply the ideal-gas law, we introduce yet another unknown, namely, temperature T. Hence, the energy equation must be solved along with the compressible forms of the equations of conservation of mass and conservation of momentum (Fig. 15–72). In addition, fluid properties, such as viscosity and thermal conductivity, are no longer necessarily treated as constants, since they are functions of temperature; thus, they appear inside the derivative operators in the differential equations of Fig. 15–72. While the equation set looks ominous, many commercially available CFD codes are able to handle compressible flow problems, including shock waves. When solving compressible flow problems with CFD, the boundary conditions are somewhat different than those of incompressible flow. For example, at a pressure inlet we need to specify both stagnation pressure and static pressure, along with stagnation temperature. A special boundary condition (called pressure far field in FLUENT) is also available for compressible flows. With this boundary condition, we specify the Mach number, the static pressure, and the temperature; it can be applied to both inlets and outlets and is well-suited for supersonic external flows. The equations of Fig. 15–72 are for laminar flow, whereas many compressible flow problems occur at high flow speeds in which the flow is turbulent. Therefore, the equations of Fig. 15–72 must be modified accordingly (into the RANS equation set) to include a turbulence model, and more transport equations must be added, as discussed previously. The equations then get quite long and complicated and are not included here. Fortunately, in many situations, we can approximate the flow as inviscid, eliminating the viscous terms from the equations of Fig. 15–72 (the Navier–Stokes equation reduces to the Euler equation). As we shall see, the inviscid flow approximation turns out to be quite good for many practical high-speed flows, since the boundary layers along walls are very thin at high Reynolds numbers. In
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861 CHAPTER 15
Continuity:
x-momentum: -momentum:
-momentum: y-momentum:
-momentum: z-momentum:
Energy:
∂(ru) ∂x
+
∂(rv) ∂y
+
∂(rw) ∂z
=0
Ideal gas law: P = rRT RT ∂w
r Qu
→ → ∂u ∂u ∂ ∂ ∂v ∂ ∂u ∂u ∂P ∂u +v + w Q = rgx – + l- • V Q + + + + 2m m ∂x ∂y ∂x Q ∂x ∂y c Q ∂y ∂x Qc ∂z ∂z ∂x
r Qu
→ → ∂w ∂ ∂w ∂ ∂w ∂ ∂v ∂w ∂P ∂w ∂u ∂w +v + w Q = rgz – + c m Q ∂x + ∂z Q c + ∂y cm Q ∂z + ∂y Qc + ∂z Q2m ∂z + l- • V Q ∂x ∂y ∂x ∂z ∂z
r Qu
→ → ∂v ∂v ∂ ∂u ∂ ∂ ∂v ∂P ∂v ∂v +v + w Q = rgy – + c m Q ∂x + ∂y Q c + ∂y Q2m ∂y + l- • V Q + ∂z ∂x ∂y ∂x ∂y ∂z
rcp Qu
cm Q ∂x
∂v
cm Q ∂z
+
∂u ∂z Qc
+
∂w Q ∂y c
→ → ∂T ∂T ∂P ∂T ∂P ∂P +v + w Q = bT Qu +v + w Q + - • (k-T ) + . ∂x ∂y ∂y ∂z ∂x ∂z
FIGURE 15–72 The equations of motion for the case of steady, compressible, laminar flow of a Newtonian fluid in Cartesian coordinates. There are six equations and six unknowns: r, u, v, w, T, and P. Five of the equations are nonlinear partial differential equations, while the ideal-gas law is an algebraic equation. R is the specific ideal-gas constant, l is the second coefficient of viscosity, often set equal to #2m/3; cp is the specific heat at constant pressure; k is the thermal conductivity; b is the coefficient of thermal expansion, and . is the dissipation function, given by White (1991) as . " 2m a
/u 2 /v 2 /w 2 /v /u 2 /w /v 2 /u /w 2 /u /v /w 2 b % 2m a b % 2m a b % m a % b % m a % b % ma % b % la % % b /x /y /z /x /y /y /z /z /x /x /y /z
fact, compressible CFD calculations can predict flow features that are often quite difficult to obtain experimentally. For example, many experimental measurement techniques require optical access, which is limited in threedimensional flows, and even in some axisymmetric flows. CFD is not limited in this way.
Compressible Flow through a Converging–Diverging Nozzle
For our first example, we consider compressible flow of air through an axisymmetric converging–diverging nozzle. The computational domain is shown in Fig. 15–73. The inlet radius is 0.10 m, the throat radius is 0.075 m, and the outlet radius is 0.12 m. The axial distance from the inlet to the throat is 0.30 m—the same as the axial distance from the throat to the outlet. A structured grid with approximately 12,000 quadrilateral cells is used in the calculations. At the pressure inlet boundary, the stagnation pressure P0, inlet is set to 220 kPa (absolute), the static pressure Pinlet is set to 210 kPa, and the stagnation temperature T0, inlet is set to 300 K. For the first case, we set the static pressure Pb at the pressure outlet boundary to 50.0 kPa (Pb /P0, inlet " 0.227)—low enough that the flow is supersonic through the entire diverging section of the nozzle, without any normal shocks in the
Pressure inlet
Pressure outlet Wall
Axis
FIGURE 15–73 Computational domain for compressible flow through a converging–diverging nozzle. Since the flow is axisymmetric, only one two-dimensional slice is needed for the CFD solution.
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862 FLUID MECHANICS
nozzle. This back pressure ratio corresponds to a value between cases E and F in Fig. 12–27, in which a complex shock pattern occurs downstream of the nozzle exit; these shock waves do not influence the flow in the nozzle itself, since the flow exiting the nozzle is supersonic. We do not attempt to model the flow downstream of the nozzle exit. The CFD code is run to convergence in its steady, inviscid, compressible flow mode. The average values of the Mach number Ma and pressure ratio P/P0, inlet are calculated at 25 axial locations along the converging–diverging nozzle (every 0.025 m) and are plotted in Fig. 15–74a. The results match almost perfectly with the predictions of one-dimensional isentropic flow (Chap. 12). At the throat (x " 0.30 m), the average Mach number is 0.997, and the average value of P/P0, inlet is 0.530. One-dimensional isentropic flow theory predicts Ma " 1 and P/P0, inlet " 0.528 at the throat. The small discrepancies between CFD and theory are due to the fact that the computed flow is not one-dimensional, since there is a radial velocity component and, therefore, a radial variation of the Mach number and static pressure. Careful examination of the Mach number contour lines of Fig. 15–74b reveal that they are curved, not straight as would be predicted by one-dimensional isentropic
1
2.5
0.9 2.0
0.8
P/P0, inlet
0.7 0.6
1.5 Ma
0.5 0.4
1.0
FIGURE 15–74 CFD results for steady, adiabatic, inviscid compressible flow through an axisymmetric converging–diverging nozzle: (a) calculated average Mach number and pressure ratio at 25 axial locations (circles), compared to predictions from isentropic, onedimensional compressible flow theory (solid lines); (b) grayscale Mach number contours, ranging from Ma " 0.3 (darkest) to 2.7 (lightest). Although only the top half is calculated, a mirror image about the x-axis is shown for clarity. The sonic line (Ma " 1) is also highlighted. It is parabolic rather than straight in this axisymmetric flow due to the radial component of velocity, as discussed in Schreier (1982).
0.3
Ma 0.5
0.2
Throat
0.1 0.0
0 0
0.1
0.2
0.3 x, m (a)
0.4
0.5
Sonic line
(b)
0.6
P P0, inlet
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863 CHAPTER 15
theory. The sonic line (Ma " 1) is identified for clarity in the figure. Although Ma " 1 right at the wall of the throat, sonic conditions along the axis of the nozzle are not reached until somewhat downstream of the throat. Next, we run a series of cases in which back pressure Pb is varied, while keeping all other boundary conditions fixed. Results for three cases are shown in Fig. 15–75: Pb " (a) 100, (b) 150, and (c) 200 kPa, i.e., Pb /P0, inlet " (a) 0.455, (b) 0.682, and (c) 0.909, respectively. For all three cases, a normal shock occurs in the diverging portion of the nozzle. Furthermore, as back pressure increases, the shock moves upstream toward the throat, and decreases in strength. Since the flow is choked at the throat, the mass flow rate is identical in all three cases (and also in the previous case shown in Fig. 15–74). We notice that the normal shock is not straight, but rather is curved due to the radial component of velocity, as previously mentioned. For case (b), in which Pb /P0, inlet " 0.682, the average values of the Mach number and pressure ratio P/P0, inlet are calculated at 25 axial locations along the converging–diverging nozzle (every 0.025 m), and are plotted in Fig. 15–76. For comparison with theory, the one-dimensional isentropic flow relations are used upstream and downstream of the shock, and the normal
Shock
(a)
Shock
(b) Shock
(c)
FIGURE 15–75 CFD results for steady, adiabatic, inviscid, compressible flow through a converging-diverging nozzle: grayscale contours of stagnation pressure ratio P0 /P0, inlet are shown for Pb /P0, inlet " (a) 0.455; (b) 0.682; and (c) 0.909. Since stagnation pressure is constant upstream of the shock and decreases suddenly across the shock, it serves as a convenient indicator of the location and strength of the normal shock in the nozzle. In these contour plots, P0 /P0, inlet ranges from 0.5 (darkest) to 1.05 (lightest). It is clear from the grayscale levels downstream of the shock that the further downstream the shock, the stronger the shock (larger magnitude of stagnation pressure drop across the shock). We also note the shape of the shocks—curved rather than straight, because of the radial component of velocity.
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864 FLUID MECHANICS 1
2.5
0.9 2.0
0.8
P/P0, inlet
0.7 0.6
1.5
FIGURE 15–76 Mach number and pressure ratio as functions of axial distance along a converging–diverging nozzle for the case in which Pb /P0, inlet " 0.682. Averaged CFD results at 25 axial locations (circles) for steady, inviscid, adiabatic, compressible flow are compared to predictions from onedimensional compressible flow theory (solid lines).
Ma
0.5
Shock
P P0, inlet
0.4
1.0
0.3
Ma 0.5
0.2 Throat
0.1
0.0
0 0
0.1
0.2
0.3
0.4
0.5
0.6
x, m
shock relations are used to calculate the pressure jump across the shock (Chap. 12). To match the specified back pressure, one-dimensional analysis requires that the normal shock be located at x " 0.4436 m, accounting for the change in both P0 and A* across the shock. The agreement between CFD calculations and one-dimensional theory is again excellent. The small discrepancy in both the pressure and the Mach number just downstream of the shock is attributed to the curved shape of the shock (Fig. 15–75b), as discussed previously. In addition, the shock in the CFD calculations is not infinitesimally thin, as predicted by one-dimensional theory, but is spread out over a few computational cells. The latter inaccuracy can be reduced somewhat by refining the grid in the area of the shock wave (not shown). The previous CFD calculations are for steady, inviscid, adiabatic flow. When there are no shock waves (Fig. 15–74), the flow is also isentropic, since it is both adiabatic and reversible (no irreversible losses). However, when a shock wave exists in the flow field (Fig. 15–75), the flow is no longer isentropic since there are irreversible losses across the shock, although it is still adiabatic. One final CFD case is run in which two additional irreversibilities are included, namely, friction and turbulence. We modify case (b) of Fig. 15–75 by running a steady, adiabatic, turbulent case using the k-e turbulence model with wall functions. The turbulence intensity at the inlet is set to 10 percent with a turbulence length scale of 0.050 m. A contour plot of P/P0, inlet is shown in Fig. 15–77, using the same grayscale range as in Fig. 15–75. Comparison of Figs. 15–75b and 15–77 reveals that the shock wave for the turbulent case occurs further upstream and is therefore somewhat weaker. In addition, the stagnation pressure is small in a very thin region along the channel walls. This is due to frictional losses in the thin boundary layer. Turbulent and viscous irreversibilities in the boundary layer region are responsible for this decrease in stagnation pressure. Furthermore, the boundary layer separates just downstream of the shock, leading to more
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865 CHAPTER 15 Shock
Boundary layer irreversibilities
Flow separation
irreversibilities. A close-up view of velocity vectors in the vicinity of the separation point along the wall is shown in Fig. 15–78. We note that this case does not converge well and is inherently unsteady; the interaction between shock waves and boundary layers is a very difficult task for CFD. Because we use wall functions, flow details within the turbulent boundary layer are not resolved in this CFD calculation. Experiments reveal, however, that the shock wave interacts much more significantly with the boundary layer, producing “l-feet,” as discussed in the Application Spotlight of Chap. 12. Finally, we compare the mass flow rate for this viscous, turbulent case to . that of the inviscid case, and find that m has decreased by about 0.7 percent. Why? As discussed in Chap. 10, a boundary layer along a wall impacts the outer flow such that the wall appears to be thicker by an amount equal to the displacement thickness d*. The effective throat area is thus reduced somewhat by the presence of the boundary layer, leading to a reduction in mass flow rate through the converging–diverging nozzle. The effect is small in this example since the boundary layers are so thin relative to the dimensions of the nozzle, and it turns out that the inviscid approximation is quite good (less than one percent error).
FIGURE 15–77 CFD results for stationary, adiabatic, turbulent, compressible flow through a converging–diverging nozzle. Grayscale contours of stagnation pressure ratio P0 /P0,inlet are shown for the case with Pb /P0,inlet " 0.682, the same back pressure as that of Fig. 15–75b. Flow separation and irreversibilities in the boundary layer are identified.
Approximate location of shock
FIGURE 15–78 Close-up view of velocity vectors in the vicinity of the separated flow region of Fig. 15–77. The sudden decrease in velocity magnitude across the shock is seen, as is the reverse flow region downstream of the shock.
Oblique Shocks over a Wedge
For our final compressible flow example, we model steady, adiabatic, twodimensional, inviscid, compressible flow of air over a wedge of half-angle u (Fig. 15–79). Since the flow has top–bottom symmetry, we model only the upper half of the flow and use a symmetry boundary condition along the bottom edge. We run three cases: u " 10, 20, and 30°, at an inlet Mach number of 2.0. CFD results are shown in Fig. 15–80 for all three cases. In the CFD plots, a mirror image of the computational domain is projected across the line of symmetry for clarity. For the 10° case (Fig. 15–80a), a straight oblique shock originating at the apex of the wedge is observed, as also predicted by inviscid theory. The flow turns across the oblique shock by 10° so that it is parallel to the wedge wall. The shock angle b predicted by inviscid theory is 39.31°, and the predicted Mach number downstream of the shock is 1.64. Measurements with a protractor on Fig. 15–80a yield b ! 40°, and the CFD calculation of the Mach number downstream of the shock is 1.64; the agreement with theory is thus excellent.
Pressure far field y
Symmetry
Wedge (wall)
u
x
FIGURE 15–79 Computational domain and boundary conditions for compressible flow over a wedge of half-angle u. Since the flow is symmetric about the x-axis, only the upper half is modeled in the CFD analysis.
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866 FLUID MECHANICS
Ma2
Ma1
b
Ma2
Ma1 b
u =10°
Oblique shock
u = 20°
Oblique shock
u = 30° Ma1
Ma2
Detached shock
(a)
(b)
(c)
FIGURE 15–80 CFD results (grayscale Mach number contours) for steady, adiabatic, inviscid, compressible flow at Ma1 " 2.0 over a wedge of half-angle u " (a) 10(, (b) 20(, and (c) 30(. The Mach number contours range from Ma " 0.2 (darkest) to 2.0 (lightest) in all cases. For the two smaller wedge half-angles, an attached weak oblique shock forms at the leading edge of the wedge, but for the 30( case, a detached shock (bow wave) forms ahead of the wedge. Shock strength increases with u, as indicated by the darker shade of gray downstream of the shock as u increases.
For the 20° case (Fig. 15–80b), the CFD calculations yield a Mach number of 1.21 downstream of the shock. The shock angle measured from the CFD calculations is about 54°. Inviscid theory predicts a Mach number of 1.21 and a shock angle of 53.4°, so again the agreement between theory and CFD is excellent. Since the shock for the 20° case is at a steeper angle (closer to a normal shock), it is stronger than the shock for the 10° case, as indicated by the darker coloring in the Mach contours downstream of the shock for the 20° case. At Mach number 2.0 in air, inviscid theory predicts that a straight oblique shock can form up to a maximum wedge half-angle of about 23° (Chap. 12). At wedge half-angles greater than this, the shock must move upstream of the wedge (become detached), forming a detached shock, which takes the shape of a bow wave (Chap. 12). The CFD results at u " 30° (Fig. 15–80c) show that this is indeed the case. The portion of the detached shock just upstream of the leading edge is a normal shock, and thus the flow downstream of that portion of the shock is subsonic. As the shock curves backward, it becomes progressively weaker, and the Mach number downstream of the shock increases, as indicated by the lighter shade of gray.
15–6
■
OPEN-CHANNEL FLOW CFD CALCULATIONS
So far, all our examples have been for one single-phase fluid (air or water). However, many commercially available CFD codes can handle flow of a mixture of gases (e.g., carbon monoxide in air), flow with two phases of the same fluid (e.g., steam and liquid water), and even flow of two fluids of different phase (e.g., liquid water and gaseous air). The latter case is of interest in this section, namely, the flow of water with a free surface, above which is gaseous air, i.e., open-channel flow. We present here some simple examples of CFD solutions of open-channel flows.
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867 CHAPTER 15
Velocity inlet
Top of domain (inviscid wall)
Vinlet Pressure outlet
Air Water Vinlet
yinlet Channel bottom (wall)
Bump (wall)
Channel bottom (wall)
FIGURE 15–81 Computational domain for steady, incompressible, two-dimensional flow of water over a bump along the bottom of a channel, with boundary conditions identified. Two fluids are modeled in the flow field—liquid water and air above the free surface of the water. Liquid depth yinlet and inlet speed Vinlet are specified.
Flow over a Bump on the Bottom of a Channel
Consider a two-dimensional channel with a flat, horizontal bottom. At a certain location along the bottom of the channel, there is a smooth bump, 1.0 m long and 0.10 m high at its center (Fig. 15–81). The velocity inlet is split into two parts—the lower part for liquid water and the upper part for air. In the CFD calculations, the inlet velocity of both the air and the water is specified as Vinlet. The water depth at the inlet of the computational domain is specified as yinlet, but the location of the water surface in the rest of the domain is calculated. The flow is modeled as inviscid. We consider cases with both subcritical and supercritical inlets (Chap. 13). Results from the CFD calculations are shown in Fig. 15–82 for three cases for comparison. For the first case (Fig. 15–82a), yinlet is specified as 0.30 m, and Vinlet is specified as 0.50 m/s. The corresponding Froude number is calculated to be Froude number:
Fr "
Vinlet 2gyinlet
"
0.50 m/s 2(9.81 m/s2)(0.30 m)
(a)
" 0.291
Since Fr ' 1, the flow at the inlet is subcritical, and the liquid surface dips slightly above the bump (Fig. 15–82a). The flow remains subcritical downstream of the bump, and the liquid surface height slowly rises back to its prebump level. The flow is thus subcritical everywhere. For the second case (Fig. 15–82b), yinlet is specified as 0.50 m, and Vinlet is specified as 4.0 m/s. The corresponding Froude number is calculated to be 1.81. Since Fr * 1, the flow at the inlet is supercritical, and the liquid surface rises above the bump (Fig. 15–82b). Far downstream, the liquid depth returns to 0.50 m, and the average velocity returns to 4.0 m/s, yielding Fr " 1.81—the same as at the inlet. Thus, this flow is supercritical everywhere. Finally, we show results for a third case (Fig. 15–82c) in which the flow entering the channel is subcritical (yinlet " 0.50 m, Vinlet " 1.0 m/s, and Fr " 0.452). In this case, the water surface dips downward over the bump, as expected for subcritical flow. However, on the downstream side of the bump, youtlet " 0.25 m, Voutlet " 2.0 m/s, and Fr " 1.28. Thus, this flow starts subcritical, but changes to supercritical downstream of the bump. If the domain had extended further downstream, we would likely see a hydraulic jump that would bring the Froude number back below unity (subcritical).
(b)
(c)
FIGURE 15–82 CFD results for incompressible, twodimensional flow of water over a bump along the channel bottom. Phase contours are plotted, where blue indicates liquid water and white indicates gaseous air: (a) subcriticalto-subcritical, (b) supercritical-tosupercritical, and (c) subcritical-tosupercritical.
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Flow through a Sluice Gate (Hydraulic Jump)
As a final example, we consider a two-dimensional channel with a flat, horizontal bottom, but this time with a sluice gate (Fig. 15–83). The water depth at the inlet of the computational domain is specified as yinlet, and the inlet flow velocity is specified as Vinlet. The bottom of the sluice gate is a distance a from the channel bottom. The flow is modeled as inviscid. We run the CFD code with yinlet " 12.0 m and Vinlet " 0.833 m/s, yielding an inlet Froude number of Frinlet " 0.0768 (subcritical). The bottom of the sluice gate is at a " 0.125 m from the channel bottom. Results from the CFD calculations are shown in Fig. 15–84. After the water passes under the sluice gate, its average velocity increases to 12.8 m/s, and its depth decreases to y " 0.78 m. Thus, Fr " 4.63 (supercritical) downstream of the sluice gate and upstream of the hydraulic jump. Some distance downstream, we see a hydraulic jump in which the average water depth increases to y " 3.54 m, and the average water velocity decreases to 2.82 m/s. The Froude number downstream of the hydraulic jump is thus Fr " 0.478 (subcritical). We notice that the downstream water depth is significantly lower than that upstream of the sluice gate, indicating relatively large dissipation through the hydraulic jump and a corresponding decrease in the specific energy of the flow (Chap. 13). The analogy between specific energy loss through a hydraulic jump in open-channel flow and stagnation pressure loss through a shock wave in compressible flow is reinforced.
FIGURE 15–83 Computational domain for steady, incompressible, two-dimensional flow of water through a sluice gate, with boundary conditions identified. Two fluids are modeled in the flow field— liquid water, and air above the free surface of the water. Liquid depth yinlet and inlet speed Vinlet are specified.
FIGURE 15–84 CFD results for incompressible, twodimensional flow of water through a sluice gate in an open channel. Phase contours are plotted, where blue indicates liquid water and white indicates gaseous air: (a) overall view of the sluice gate and hydraulic jump, and (b) close-up view of the hydraulic jump. The flow is highly unsteady, and these are instantaneous snapshots at an arbitrary time.
Velocity inlet
Vinlet Air Water
Top of domain (inviscid wall)
Sluice gate (wall)
Pressure outlet
yinlet Vinlet
a Channel bottom (wall)
Sluice gate
Hydraulic jump
(a)
(b)
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869 CHAPTER 15
APPLICATION SPOTLIGHT
■
A Virtual Stomach
Guest Authors: James G. Brasseur and Anupam Pal, The Pennsylvania State University
References Indireshkumar, K., Brasseur, J. G., Faas, H., Hebbard, G. S., Kunz, P., Dent, J., Boesinger, P., Feinle, C., Fried, M., Li, M., and Schwizer, W., “Relative Contribution of ‘Pressure Pump’ and ‘Peristaltic Pump’ to Slowed Gastric Emptying,” Amer J Physiol, 278, pp. G604–616, 2000. Pal, A., Indireshkumar, K., Schwizer, W., Abrahamsson, B., Fried, M., Brasseur, J. G., “2004 Gastric Flow and Mixing Studied Using Computer Simulation,” In Press, Proc. Royal Soc. London, Biological Sciences, October.
Pylorus Antrum
Antral CW
FIGURE 15–85 Magnetic resonance image of the human stomach in vivo at one instant in time showing peristaltic (i.e., propagating) contraction waves (CW) in the end region of the stomach (the antrum). The pylorus is a sphincter, or valve, that allows nutrients into the duodenum (small intestines). Developed by Anupam Pal and James Brasseur. Used by permission.
15
y(cm)
The mechanical function of the stomach (called gastric “motility”) is central to proper nutrition, reliable drug delivery, and many gastric dysfunctions such as gastroparesis. Figure 15–85 shows a magnetic resonance image (MRI) of the stomach. The stomach is a mixer, a grinder, a storage chamber, and a sophisticated pump that controls the release of liquid and solid gastric content into the small intestines where nutrient uptake occurs. Nutrient release is controlled by the opening and closing of a valve at the end of the stomach (the pylorus) and the time variations in pressure difference between the stomach and duodenum. Gastric pressure is controlled by muscle tension over the stomach wall and peristaltic contraction waves that pass through the antrum (Fig. 15–85). These antral peristaltic contraction waves also break down food particles and mix material within the stomach, both food and drugs. It is currently impossible, however, to measure the mixing fluid motions in the human stomach. The MRI, for example, gives only an outline of special magnetized fluid within the stomach. In order to study these invisible fluid motions and their effects, we have developed a computer model of the stomach using computational fluid dynamics. The mathematics underlying our computational model is derived from the laws of fluid mechanics. The model is a way of extending MRI measurements of time-evolving stomach geometry to the fluid motions within. Whereas computer models cannot describe the full complexity of gastric physiology, they have the great advantage of allowing controlled systematic variation of parameters, so sensitivities that cannot be measured experimentally can be studied computationally. Our virtual stomach applies a numerical method called the “lattice Boltzmann” algorithm that is well suited to fluid flows in complex geometries, and the boundary conditions are obtained from MRI data. In Fig. 15–86 we predict the motions, breakdown, and mixing of 1-cm-size extended-release drug tablets in the stomach. In this numerical experiment the drug tablet is denser than the surrounding highly viscous meal. We predict that the antral peristaltic waves generate recirculating eddies and retropulsive “jets” within the stomach, which in turn generate high shear stresses that wear away the tablet surface and release the drug. The drug then mixes by the same fluid motions that release the drug. We find that gastric fluid motions and mixing depend on the details of the time variations in stomach geometry and pylorus.
10
5
0 0
5
10 x(cm)
15
20
FIGURE 15–86 Computer simulation of fluid motions within the stomach (velocity vectors) from peristaltic antral contraction waves (Fig. 15–85), and the release of a drug (blue trail) from an extended release tablet (blue circle). Developed by Anupam Pal and James Brasseur. Used by permission.
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SUMMARY Although neither as ubiquitous as spreadsheets, nor as easy to use as mathematical solvers, computational fluid dynamics codes are continually improving and are becoming more commonplace. Once the realm of specialized scientists who wrote their own codes and used supercomputers, commercial CFD codes with numerous features and user-friendly interfaces can now be obtained for personal computers at a reasonable cost and are available to engineers of all disciplines. As shown in this chapter, however, a poor grid, improper choice of laminar versus turbulent flow, inappropriate boundary conditions, and/or any of a number of other miscues can lead to CFD solutions that are physically incorrect, even though the colorful graphical output always looks pretty. Therefore, it is imperative that CFD users be well grounded in the fundamentals of fluid mechanics in order to avoid erroneous answers from a CFD simulation. In addition, appropriate comparisons should be made to experimental data whenever possible to validate CFD predictions. Bearing these cautions in mind, CFD has enormous potential for diverse applications involving fluid flows. We show examples of both laminar and turbulent CFD solutions. For incompressible laminar flow, computational fluid dynamics does an excellent job, even for unsteady flows with separation. In fact, laminar CFD solutions are “exact” to the extent that they are limited by grid resolution and boundary conditions. Unfortunately, many flows of practical engineering interest are turbulent, not laminar. Direct numerical simulation (DNS) has great potential for simulation of complex turbulent flow fields, and algorithms for solving the equations of motion (the three-dimensional continuity and Navier–Stokes equations) are well established. However, resolution of all the fine scales of a high Reynolds number com-
plex turbulent flow requires computers that are orders of magnitude faster than today’s fastest machines. It will be decades before computers advance to the point where DNS is useful for practical engineering problems. In the meantime, the best we can do is employ turbulence models, which are semi-empirical transport equations that model (rather than solve) the increased mixing and diffusion caused by turbulent eddies. When running CFD codes that utilize turbulence models, we must be careful that we have a fine-enough mesh and that all boundary conditions are properly applied. In the end, however, regardless of how fine the mesh, or how valid the boundary conditions, turbulent CFD results are only as good as the turbulence model used. Nevertheless, while no turbulence model is universal (applicable to all turbulent flows), we obtain reasonable performance for many practical flow simulations. We also demonstrate in this chapter that CFD can yield useful results for flows with heat transfer, compressible flows, and open-channel flows. In all cases, however, users of CFD must be careful that they choose an appropriate computational domain, apply proper boundary conditions, generate a good grid, and use the proper models and approximations. As computers continue to become faster and more powerful, CFD will take on an ever-increasing role in design and analysis of complex engineering systems. We have only scratched the surface of computational fluid dynamics in this brief chapter. In order to become proficient and competent at CFD, you must take advanced courses of study in numerical methods, fluid mechanics, turbulence, and heat transfer. We hope that, if nothing else, this chapter has spurred you on to further study of this exciting topic.
REFERENCES AND SUGGESTED READING 1. C-J. Chen and S-Y. Jaw. Fundamentals of Turbulence Modeling. Washington, DC: Taylor & Francis, 1998.
6. D. J. Tritton. Physical Fluid Dynamics. New York: Van Nostrand Reinhold Co., 1977.
2. J. M. Cimbala, H. Nagib, and A. Roshko. “Large Structure in the Far Wakes of Two-Dimensional Bluff Bodies,” Fluid Mech., 190, pp. 265–298, 1988.
7. M. Van Dyke. An Album of Fluid Motion. Stanford, CA: The Parabolic Press, 1982.
3. S. Schreier. Compressible Flow. New York: WileyInterscience, chap. 6 (Transonic Flow), pp. 285–293, 1982. 4. J. C. Tannehill, D. A. Anderson, and R. H. Pletcher. Computational Fluid Mechanics and Heat Transfer, 2nd ed. Washington, DC: Taylor & Francis, 1997. 5. H. Tennekes and J. L. Lumley. A First Course in Turbulence. Cambridge, MA: The MIT Press, 1972.
8. F. M. White. Viscous Fluid Flow, 2nd ed. New York: McGraw-Hill, 1991. 9. D. C. Wilcox. Turbulence Modeling for CFD, 2nd ed. La Cañada, CA: DCW Industries, Inc., 1998. 10. C. H. K. Williamson. “Oblique and Parallel Modes of Vortex Shedding in the Wake of a Circular Cylinder at Low Reynolds Numbers,” J. Fluid Mech., 206, pp. 579–627, 1989.
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PROBLEMS* Fundamentals, Grid Generation, and Boundary Conditions 15–1C A CFD code is used to solve a two-dimensional (x and y), incompressible, laminar flow without free surfaces. The fluid is Newtonian. Appropriate boundary conditions are used. List the variables (unknowns) in the problem, and list the corresponding equations to be solved by the computer. 15–2C Write a brief (a few sentences) definition and description of each of the following, and provide example(s) if helpful: (a) computational domain, (b) mesh, (c) transport equation, (d ) coupled equations. 15–3C What is the difference between a node and an interval and how are they related to cells? In Fig. P15–3C, how many nodes and how many intervals are on each edge?
15–7C Write a brief (a few sentences) discussion about the significance of each of the following in regards to an iterative CFD solution: (a) initial conditions, (b) residual, (c) iteration, (d) postprocessing. 15–8C Briefly discuss how each of the following is used by CFD codes to speed up the iteration process: (a) multigridding and (b) artificial time. 15–9C Of the boundary conditions discussed in this chapter, list all the boundary conditions that may be applied to the right edge of the two-dimensional computational domain sketched in Fig. P15–9C. Why can’t the other boundary conditions be applied to this edge? 15–10C What is the standard method to test for adequate grid resolution when using CFD?
BC to be specified on this edge
FIGURE P15–3C 15–4C For the two-dimensional computational domain of Fig. P15–3C, with the given node distribution, sketch a simple structured grid using four-sided cells and sketch a simple unstructured grid using three-sided cells. How many cells are in each? Discuss. 15–5C Summarize the eight steps involved in a typical CFD analysis of a steady, laminar flow field. 15–6C Suppose you are using CFD to simulate flow through a duct in which there is a circular cylinder as in Fig. P15–6C. The duct is long, but to save computer resources you choose a computational domain in the vicinity of the cylinder only. Explain why the downstream edge of the computational domain should be further from the cylinder than the upstream edge. Computational domain In
FIGURE P15–9C 15–11C What is the difference between a pressure inlet and a velocity inlet boundary condition? Explain why you cannot specify both pressure and velocity at a velocity inlet boundary condition or at a pressure inlet boundary condition. 15–12C An incompressible CFD code is used to simulate the flow of air through a two-dimensional rectangular channel (Fig. P15–12C). The computational domain consists of four blocks, as indicated. Flow enters block 4 from the upper right and exits block 1 to the left as shown. Inlet velocity V is known and outlet pressure Pout is also known. Label the boundary conditions that should be applied to every edge of every block of this computational domain.
Out Block 1 Pout
FIGURE P15–6C Out
* Problems designated by a “C” are concept questions, and students are encouraged to answer them all. Problems designated by an “E” are in English units, and the SI users can ignore them.
In
V Block 4
Block 2
Block 3
FIGURE P15–12C
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15–13C Consider Prob. 15–12C again, except let the boundary condition on the common edge between blocks 1 and 2 be a fan with a specified pressure rise from right to left across the fan. Suppose an incompressible CFD code is run for both cases (with and without the fan). All else being equal, will the pressure at the inlet increase or decrease? Why? What will happen to the velocity at the outlet? Explain. 15–14C List six boundary conditions that are used with CFD to solve incompressible fluid flow problems. For each one, provide a brief description and give an example of how that boundary condition is used. 15–15 A CFD code is used to simulate flow over a twodimensional airfoil at an angle of attack. A portion of the computational domain near the airfoil is outlined in Fig. P15–15 (the computational domain extends well beyond the region outlined by the dashed line). Sketch a coarse structured grid using four-sided cells and sketch a coarse unstructured grid using three-sided cells in the region shown. Be sure to cluster the cells where appropriate. Discuss the advantages and disadvantages of each grid type.
four-sided blocks, and sketch a coarse grid using four-sided cells, being sure to cluster cells near walls. Also be careful to avoid highly skewed cells. Label the boundary conditions that should be applied to every edge of every block of your computational domain. (Hint: Six to seven blocks are sufficient.) 15–18 An incompressible CFD code is used to simulate the flow of gasoline through a two-dimensional rectangular channel in which there is a large circular settling chamber (Fig. P15–18). Flow enters from the left and exits to the right as shown. A time-averaged turbulent flow solution is generated using a turbulence model. Top–bottom symmetry is assumed. Inlet velocity V is known, and outlet pressure Pout is also known. Generate the blocking for a structured grid using four-sided blocks, and sketch a coarse grid using four-sided cells, being sure to cluster cells near walls. Also be careful to avoid highly skewed cells. Label the boundary conditions that should be applied to every edge of every block of your computational domain.
V
In
Out
Pout
FIGURE P15–18
FIGURE P15–15 15–16 For the airfoil of Prob. 15–15, sketch a coarse hybrid grid and explain the advantages of such a grid. 15–17 An incompressible CFD code is used to simulate the flow of water through a two-dimensional rectangular channel in which there is a circular cylinder (Fig. P15–17). A timeaveraged turbulent flow solution is generated using a turbulence model. Top–bottom symmetry about the cylinder is assumed. Flow enters from the left and exits to the right as shown. Inlet velocity V is known, and outlet pressure Pout is also known. Generate the blocking for a structured grid using
V
In
Out
In
Out
FIGURE P15–17
Pout
15–19 Redraw the structured multiblock grid of Fig. 15–12b for the case in which your CFD code can handle only elementary blocks. Renumber all the blocks and indicate how many i- and j-intervals are contained in each block. How many elementary blocks do you end up with? Add up all the cells, and verify that the total number of cells does not change. 15–20 Suppose your CFD code can handle nonelementary blocks. Combine as many blocks of Fig. 15–12b as you can. The only restriction is that in any one block, the number of iintervals and the number of j-intervals must be constants. Show that you can create a structured grid with only three nonelementary blocks. Renumber all the blocks and indicate how many i- and j-intervals are contained in each block. Add up all the cells and verify that the total number of cells does not change. 15–21 A new heat exchanger is being designed with the goal of mixing the fluid downstream of each stage as thoroughly as possible. Anita comes up with a design whose cross section for one stage is sketched in Fig. P15–21. The geometry extends periodically up and down beyond the region shown here. She uses several dozen rectangular tubes inclined at a high angle of attack to ensure that the flow separates and
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mixes in the wakes. The performance of this geometry is to be tested using two-dimensional time-averaged CFD simulations with a turbulence model, and the results will be compared to those of competing geometries. Sketch the simplest possible computational domain that can be used to simulate this flow. Label and indicate all boundary conditions on your diagram. Discuss.
of a preliminary single-stage CFD analysis. Now she is asked to simulate two stages of the heat exchanger. The second row of rectangular tubes is staggered and inclined oppositely to that of the first row to promote mixing (Fig. P15–24). The geometry extends periodically up and down beyond the region shown here. Sketch a computational domain that can be used to simulate this flow. Label and indicate all boundary conditions on your diagram. Discuss. 15–25 Sketch a structured multiblock grid with four-sided elementary blocks for the computational domain of Prob. 15–24. Each block is to have four-sided structured cells, but you do not have to sketch the grid, just the block topology. Try to make all the blocks as rectangular as possible to avoid highly skewed cells in the corners. Assume that the CFD code requires that the node distribution on periodic pairs of edges be identical (the two edges of a periodic pair are “linked” in the grid generation process). Also assume that the CFD code does not allow a block’s edges to be split for application of boundary conditions.
FlowLab Problems*
FIGURE P15–21 15–22 Sketch a coarse structured multiblock grid with foursided elementary blocks and four-sided cells for the computational domain of Prob. 15–21. 15–23 Anita runs a CFD code using the computational domain and grid developed in Probs. 15–21 and 15–22. Unfortunately, the CFD code has a difficult time converging and Anita realizes that there is reverse flow at the outlet (far right edge of the computational domain). Explain why there is reverse flow, and discuss what Anita should do to correct the problem. 15–24 As a follow-up to the heat exchanger design of Prob. 15–21, suppose Anita’s design is chosen based on the results
FIGURE P15–24
15–26 In this exercise, we examine how far away the boundary of the computational domain needs to be when simulating external flow around a body in a free stream. We choose a two-dimensional case for simplicity—flow at speed V over a rectangular block whose length L is 1.5 times its height D (Fig. P15–26a). We assume the flow to be symmetric about the centerline (x-axis), so that we need to model only the upper half of the flow. We set up a semicircular computational domain for the CFD solution, as sketched in Fig. P15–26b. Boundary conditions are shown on all edges. We run several values of outer edge radius R (5 ' R/D ' 500) to determine when the far field boundary is “far enough” away. Run FlowLab, and start template Block_domain. (a) Calculate the Reynolds number based on the block height D. What is the experimentally measured value of the drag coefficient for this two-dimensional block at this Reynolds number (see Chap. 11)? (b) Generate CFD solutions for various values of R/D. For each case, calculate and record drag coefficient CD. Plot CD as a function of R/D. At what value of R/D does CD become independent of computational extent to three significant digits of precision? Report a final value of CD, and discuss your results.
* These problems require the CFD software program FlowLab, provided with this textbook by FLUENT, Inc. Templates for these problems are available on the book’s website. In each case, a brief statement of the problem is provided here, whereas additional details about the geometry, boundary conditions, and computational parameters are provided within the template.
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Inlet Block V
y x
D
Outlet
V y
Wall
Wall
L
R
x
Block Symmetry (a)
Symmetry (b)
FIGURE P15–26 (c) Discuss some reasons for the discrepancy between the experimental value of CD and the value obtained here using CFD. (d) Plot streamlines for two cases: R/D " 5 and 500. Compare and discuss. 15–27 Using the geometry of Prob. 15–26, and the case with R/D " 500, the goal of this exercise is to check for grid independence. Run FlowLab, and start template Block_mesh. Run various values of grid resolution, and tabulate drag coefficient CD as a function of the number of cells. Has grid independence been achieved? Report a final value of CD to three significant digits of precision. Does the final value of drag coefficient agree better with that of this experiment? Discuss. 15–28 In Probs. 15–26 and 15–27, we used air as the fluid in our calculations. In this exercise, we repeat the calculation of drag coefficient, except we use different fluids. We adjust the inlet velocity appropriately such that the calculations are always at the same Reynolds number. Run FlowLab, and start template Block_ fluid. Compare the value of CD for all three cases (air, water, and kerosene) and discuss. 15–29 Experiments on two-dimensional rectangular blocks in an incompressible free-stream flow reveal that the drag coefficient is independent of Reynolds number for Re greater than about 104. In this exercise, we examine if CFD calculations are able to predict the same independence of CD on Re. Run FlowLab, and start template Block_Reynolds. Calculate and record CD for several values of Re. Discuss. 15–30 In Probs. 15–26 through 15–29, the k-e turbulence model is used. The goal of this exercise is to see how sensitive the drag coefficient is to our choice of turbulence model and to see if a different turbulence model yields better agreement with experiment. Run FlowLab, and start template Block_turbulence_model. Run the simulation with all the available turbulence models. For each case, record CD. Which one gives the best agreement with experiment? Discuss. 15–31 Experimental drag coefficient data are available for two-dimensional blocks of various shapes in external flow. In this exercise, we use CFD to compare the drag coefficient of
rectangular blocks with L/D ranging from 0.1 to 3.0 (Fig. P15–31). The computational domain is a semicircle similar to that sketched in Fig. P15–26b; we assume steady, incompressible, turbulent flow with symmetry about the x-axis. Run FlowLab, and start template Block_length. (a) Run the CFD simulation for various values of L/D between 0.1 and 3.0. Record the drag coefficient for each case, and plot CD as a function of L/D. Compare to experimentally obtained data on the same plot. Discuss. (b) For each case, plot streamlines near the block and in its wake region. Use these streamlines to help explain the trend in the plot of CD versus L/D. (c) Discuss possible reasons for the discrepancy between CFD calculations and experimental data and suggest a remedy. y V
x
D Block L y
V
D
x
Block L
FIGURE P15–31 15–32 Repeat Prob. 15–26 for the case of axisymmetric flow over a blunt-faced cylinder (Fig. P15–32), using FlowLab x
V
D L
FIGURE P15–32
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Wall
Wall
V Pout
D2
D1 x
L1
L2
Pin V
(a)
x
Axis
(b)
FIGURE P15–33
template Block_axisymmetric. The grids and all the parameters are the same as those in Prob. 15–26, except the symmetry boundary condition is changed to “axis,” and the flow solver is axisymmetric about the x-axis. In addition to the questions listed in Prob. 15–26, compare the two-dimensional and axisymmetric cases. Which one requires a greater extent of the far field boundary? Which one has better agreement with experiment? Discuss. (Note: The reference area for CD in the axisymmetric case is the frontal area A " pD2/4.) 15–33 Air flows through a conical diffuser in an axisymmetric wind tunnel (Fig. P15–33a—drawing not to scale). u is the diffuser half-angle (the total angle of the diffuser is equal to 2u). The inlet and outlet diameters are D1 " 0.50 m and D2 " 1.0 m, respectively, and u " 20°. The inlet velocity is nearly uniform at V " 10.0 m/s. The axial distance upstream of the diffuser is L1 " 1.50 m, and the axial distance from the start of the diffuser to the outlet is L2 " 8.00 m. We set up a computational domain for a CFD solution, as sketched in Fig. P15–33b. Since the flow is axisymmetric and steady in the mean, we model only one two-dimensional slice as shown, with the bottom edge of the domain specified as an axis. The goal of this exercise is to test for grid independence. Run FlowLab, and start template Diffuser_mesh. (a) Generate CFD solutions for several grid resolutions. Plot streamlines in the diffuser section for each case. At what grid resolution does the streamline pattern appear to be grid independent? Describe the flow field for each case and discuss. (b) For each case, calculate and record pressure difference )P " Pin # Pout. At what grid resolution is the )P grid independent (to three significant digits of precision)? Plot )P as a function of number of cells. Discuss your results. 15–34 Repeat Prob. 15–33 for the finest resolution case, but with the “pressure outlet” boundary condition changed to an “outflow” boundary condition instead, using FlowLab template Diffuser_outflow. Record )P and compare with the result of Prob. 15–33 for the same grid resolution. Also compare the pressure distribution at the outlet for the case with the pressure outlet boundary condition and the case with the outflow boundary condition. Discuss.
15–35 Barbara is designing a conical diffuser for the axisymmetric wind tunnel of Prob. 15–33. She needs to achieve at least 40 Pa of pressure recovery through the diffuser, while keeping the diffuser length as small as possible. Barbara decides to use CFD to compare the performance of diffusers of various half-angles (5° 0 u 0 90°) (see Fig. P15–33 for the definition of u and other parameters in the problem). In all cases, the diameter doubles through the diffuser—the inlet and outlet diameters are D1 " 0.50 m and D2 " 1.0 m, respectively. The inlet velocity is nearly uniform at V " 10.0 m/s. The axial distance upstream of the diffuser is L1 " 1.50 m, and the axial distance from the start of the diffuser to the outlet is L2 " 8.00 m. (The overall length of the computational domain is 9.50 m in all cases.) Run FlowLab with template Diffuser_angle. In addition to the axis and wall boundary conditions labeled in Fig. P15–33, the inlet is specified as a velocity inlet and the outlet is specified as a pressure outlet with Pout " 0 gage pressure for all cases. The fluid is air at default conditions, and turbulent flow is assumed. (a) Generate CFD solutions for half-angle u " 5, 7.5, 10, 12.5, 15, 17.5, 20, 25, 30, 45, 60, and 90°. Plot streamlines for each case. Describe how the flow field changes with the diffuser half-angle, paying particular attention to flow separation on the diffuser wall. How small must u be to avoid flow separation? (b) For each case, calculate and record )P " Pin # Pout. Plot )P as a function of u and discuss your results. What is the maximum value of u that achieves Barbara’s design objectives? 15–36 Consider the diffuser of Prob. 15–35 with u " 90° (sudden expansion). In this exercise, we test whether the grid is fine enough by performing a grid independence check. Run FlowLab, and start template Expansion_mesh. Run the CFD code for several levels of grid refinement. Calculate and record )P for each case. Discuss. 15–37 Water flows through a sudden contraction in a small round tube (Fig. P15–37a). The tube diameters are D1 " 8.0 mm and D2 " 2.0 mm. The inlet velocity is nearly uniform at V " 0.050 m/s, and the flow is laminar. Shane wants to predict the pressure difference from the inlet (x " #L1) to
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876 FLUID MECHANICS Wall V
D1 V
D2
Wall
Pin
Pout P1 Axis
x Lextend
x
L1 (a)
(b)
FIGURE P15–37
the axial location of the sudden contraction (x " 0). He sets up the computational domain sketched in Fig. P15–37b. Since the flow is axisymmetric and steady, Shane models only one slice, as shown, with the bottom edge of the domain specified as an axis. In addition to the boundary conditions labeled in Fig. P15–37b, the inlet is specified as a velocity inlet, and the outlet is specified as a pressure outlet with Pout " 0 gage. What Shane does not know is how far he needs to extend the domain downstream of the contraction in order for the flow field to be simulated accurately upstream of the contraction. (He has no interest in the flow downstream of the contraction.) In other words, he does not know how long to make Lextend. Run FlowLab, and start template Contraction_domain. (a) Generate solutions for Lextend/D2 " 0.25, 0.5, 0.75, 1.0, 1.25, 1.5, 2.0, 2.5, and 3.0. How big must Lextend/D2 be in order to avoid reverse flow at the pressure outlet? Explain. Plot streamlines near the sudden contraction to help explain your results. (b) For each case, record gage pressures Pin and P1, and calculate )P " Pin # P1. How big must Lextend/D2 be in order for )P to become independent of Lextend (to three significant digits of precision)? (c) Plot inlet gage pressure Pin as a function of Lextend/D2. Discuss and explain the trend. Based on all your results taken collectively, which value of Lextend/D2 would you recommend to Shane? 15–38 Consider the sudden contraction of Prob. 15–37 (Fig. P15–37). Suppose Shane were to disregard the downstream extension entirely (Lextend/D2 " 0). Run FlowLab, and start template Contraction_zerolength. Iterate to convergence. Is there reverse flow? Explain. Plot streamlines near the outlet, and compare with those of Prob. 15–37. Discuss. Calculate )P " Pin # Pout, and calculate the percentage error in )P under these conditions, compared to the converged value of Prob. 15–37. Discuss. 15–39 In this exercise, we apply different back pressures to the sudden contraction of Prob. 15–37 (Fig. P15–37), for the case with Lextend/D2 " 2.0. Run FlowLab, and start template Contraction_pressure. Set the pressure boundary condition at
the outlet to Pout " #50,000 Pa gage (about 1/2 atm below atmospheric pressure). Record Pin and P1, and calculate )P " Pin # P1. Repeat for Pout " 0 Pa gage and Pout " 50,000 Pa gage. Discuss your results. 15–40 Consider the sudden contraction of Prob. 15–37, but this time with turbulent rather than laminar flow. The dimensions shown in Fig. P15–37 are scaled proportionally by a factor of 100 everywhere so that D1 " 0.80 m and D2 " 0.20 m. The inlet velocity is also increased to V " 1.0 m/s. A 10 percent turbulence intensity is specified at the inlet. The outlet pressure is fixed at zero gage pressure for all cases. Run FlowLab, and start template Contraction_turbulent. (a) Calculate the Reynolds numbers of flow through the large tube and the small tube for Prob. 15–37 and also for this problem. Are our assumptions of laminar versus turbulent flow reasonable for these problems? (b) Generate CFD solutions for Lextend/D2 " 0.25, 0.5, 0.75, 1.0, 1.25, 1.5, and 2.0. How big must Lextend/D2 be in order to avoid reverse flow at the pressure outlet? Plot streamlines for the case in which Lextend/D2 " 0.75 and compare to the corresponding streamlines of Prob. 15–37 (laminar flow). Discuss. (c) For each case, record gage pressures Pin and P1, and calculate )P " Pin # P1. How big must Lextend/D2 be in order for )P to become independent of Lextend (to three significant digits of precision)? 15–41 Run FlowLab, and start template Contraction_outflow. The conditions are identical to Prob. 15–40 for the case with Lextend/D2 " 0.75, but with the “pressure outlet” boundary condition changed to an “outflow” boundary condition instead. Record Pin and P1, calculate )P " Pin # P1, and compare with the result of Prob. 15–40 for the same geometry. Discuss. 15–42 Run FlowLab with template Contraction_2d. This is identical to the sudden contraction of Prob. 15–40, but the flow is two-dimensional instead of axisymmetric. (Note that the “axis” boundary condition is replaced by “symmetry.”) As previously, the outlet pressure is set to zero gage pressure. (a) Generate CFD solutions for Lextend/D2 " 0.25, 0.5, 0.75, 1.0, 1.25, 1.5, 2.0, 3.0, and 4.0. How big must Lextend/D2 be in
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877 CHAPTER 15 y
L1
V
Wall
V Lj
Pout
x
D1
Pin L2
Wall
(a)
Wall (b)
FIGURE P15–43 order to avoid reverse flow at the pressure outlet? Plot streamlines for the case in which Lextend/D2 " 0.75, and compare to the corresponding streamlines of Prob. 15–40 (axisymmetric flow). Discuss. (b) For each case, record gage pressures Pin and P1, and calculate )P " Pin # P1. How big must Lextend/D2 be in order for )P to become independent of Lextend (to three significant digits of precision)? 15–43 Air flows through a “jog” in a rectangular channel (Fig. P15–43a, not to scale). The channel dimension is D1 " 1.0 m everywhere, and it is wide enough (into the page of Fig. P15–43) that the flow can be considered two-dimensional. The inlet velocity is nearly uniform at V " 1.0 m/s. The distance upstream of the jog is L1 " 5.0 m, the overall jog length is Lj " 3.0 m, and the distance from the end of the jog to the outlet is L2 " 10.0 m. We set up a computational domain for a CFD solution, as sketched in Fig. P15–43b. In addition to the wall boundary conditions labeled in Fig. P15–43b, the inlet is specified as a velocity inlet and the outlet is specified as a pressure outlet with Pout " 0 gage pressure. The fluid is air at default conditions, and turbulent flow is assumed. The goal of this exercise is to test for grid independence in this flow field. Run FlowLab with template Jog_turbulent_mesh. (a) Generate CFD solutions for various levels of grid resolution. Plot streamlines in the region of the jog for each case. At what grid resolution does the streamline pattern appear to be grid independent? Discuss. (b) For each case, calculate and record )P " Pin # Pout. At what grid resolution is )P grid independent (to three significant digits of precision)? Plot )P as a function of the number of cells. Discuss your results. 15–44 Repeat Prob. 15–43, but for laminar flow, using Jog_laminar_mesh as the FlowLab template. The jog is identical in shape, but scaled down by a factor of 1000 compared to that of Prob. 15–43 (the channel width is D1 " 1.0 mm everywhere). The inlet velocity is nearly uniform at V " 0.10 m/s, and the fluid is changed to water at room temperature. Discuss your results. 15–45 Repeat Prob. 15–44, but for laminar flow at a higher Reynolds number, using FlowLab template Jog_high_Re.
Everything is identical to Prob. 15–44, except the inlet velocity is increased from to V " 0.10 to 1.0 m/s. Compare results and the Reynolds numbers for the two cases and discuss. 15–46 Consider compressible flow of air through an axisymmetric converging–diverging nozzle (Fig. P15–46), in which the inviscid flow approximation is applied. The inlet conditions are fixed (P0, inlet " 220 kPa, Pinlet " 210 kPa, and T0, inlet " 300 K), but the back pressure Pb can be varied. Run FlowLab, using template Nozzle_axisymmetric. Do several cases, with back pressure ranging from 100 to 219 kPa. For each case, calculate the mass flow rate (kg/s) through the . nozzle, and plot m as a function of Pb /P0, inlet. Explain your results. Pressure inlet
Pressure outlet Wall
Axis
FIGURE P15–46 15–47 Run FlowLab with template Nozzle_axisymmetric (Prob. 15–46). For the case in which Pb " 100 kPa (Pb /P0, inlet " 0.455), plot pressure and the Mach number contours to verify that a normal shock is present near the outlet of the computational domain. Generate a plot of average Mach number Ma and average pressure ratio P/P0, inlet across several cross sections of the domain, as in Fig. 15–76. Point out the location of the normal shock, and compare the CFD results to one-dimensional compressible flow theory. Repeat for Pb " 215 kPa (Pb /P0, inlet " 0.977). Explain. 15–48 Run FlowLab with template Nozzle_2d, which is the same as Prob. 15–46, except the flow is two-dimensional instead of axisymmetric. Note that the “axis” boundary condition is also changed to “symmetry.” Compare your results and discuss the similarities and differences. 15–49 Consider flow over a simplified, two-dimensional model of an automobile (Fig. P15–49). The inlet conditions
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are fixed at V " 60.0 mi/h (26.8 m/s), with 10 percent turbulence intensity. The standard k-e turbulence model is used. Run FlowLab with template Automobile_drag. Vary the shape of the rear end of the car, and record the drag coefficient for each shape. Also plot velocity vectors in the vicinity of the rear end for each case. Compare and discuss. Which case gives the lowest drag coefficient? Why?
Symmetry
the flow at various values of the Reynolds number Re, where Re is based on pipe diameter and average speed through the pipe. For each case, study the velocity profiles at several axial locations down the pipe, and estimate the entrance length in each case. Also plot the pressure distribution along the pipe axis for each case. Estimate the end of the entrance region as the location where the pressure begins to drop linearly with x. Compare your results with those obtained from the velocity profiles, and also with theory, Le/D ! 0.06Re. Discuss.
Pressure outlet
V Velocity inlet
D V H
Model (wall) h Ground (wall)
FIGURE P15–49 15–50 In this exercise, we examine the effect of the location of the upper symmetry boundary condition of Prob. 15–49. Run FlowLab with template Automobile_domain for several values of H/h (Fig. P15–49). Plot the calculated value of CD as a function of H/h. At what value of H/h does CD level off? In other words, how far away must the upper symmetry boundary be in order to have negligible influence on the calculated value of drag coefficient? Discuss. 15–51 Run FlowLab, and start template Automobile_turbulence_model. In this exercise, we examine the effect of turbulence model on the calculation of drag on a simplified, twodimensional model of a car (Fig. P15–49). Run all the available turbulence models. For each case, record CD. Is there much variation in the calculated values of CD? Which one is correct? Discuss. 15–52 Run FlowLab, and start template Automobile_3d. In this exercise, we compare the drag coefficient for a fully three-dimensional automobile to that predicted by the twodimensional approximation of Prob. 15–49. Note that the solution takes a long time to converge and requires a significant amount of computer resources. Therefore, the converged solution is already available in this template. Observe the three-dimensional pathlines around the car by rotating the view. Calculate the drag coefficient. Is it larger or smaller than the two-dimensional prediction? Discuss. 15–53 Run FlowLab, and start template Pipe_laminar_ developing. In this exercise, we study laminar flow in the entrance region of a round pipe (Fig. P15–53, not to scale). Because of the axisymmetry, the computational domain consists of one slice (light blue region in Fig. P15–53). Calculate
x=0
x=L
FIGURE P15–53 15–54 Run FlowLab, and start template Pipe_turbulent_ developing. In this exercise, we study turbulent flow in the entrance region of a round pipe (Fig. P15–53). Calculate the flow at several values of the Reynolds number. For each case, study the velocity profiles at several axial locations down the pipe, and estimate the entrance length in each case. Also plot the pressure distribution along the pipe axis for each case. Estimate the end of the entrance region as the location where the pressure begins to drop linearly with x. Compare your results with those obtained from the velocity profiles, and also with the empirical approximation, Le /D ! 4.4Re1/6. Compare your results to those of the laminar flow of Prob. 15–53. Discuss. Which flow regime, laminar or turbulent, has the longer entrance length? Why? 15–55 Consider fully developed, laminar pipe flow (Fig. P15–55). In this exercise, we are not concerned about entrance effects. Instead, we want to analyze the fully developed flow downstream of the entrance region. Because of the axisymmetry, the computational domain consists of one slice (light blue region). The velocity profile at the inlet boundary is set to be the same as that at the outlet boundary, but a pressure drop from x " 0 to L is imposed to simulate fully developed flow. Run FlowLab with template Pipe_laminar_developed. The template is set up such that the outlet velocity profile gets fed into the inlet. In other words, the inlet and outlet are periodic Inlet
x=0
FIGURE P15–55
D
Outlet
x=L
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boundary conditions, but with an imposed pressure drop. Run several cases corresponding to various values of the Reynolds number. For each case, look at velocity profiles to confirm that the flow is fully developed. Calculate and plot Darcy friction factor f as a function of Re, and compare with the theoretical value for laminar flow, f " 64/Re. Discuss the agreement between CFD and theory.
15–60 Run FlowLab, and start template Plate_turbulence_models. In this exercise, we examine the effect of the turbulence model on the calculation of the drag coefficient on a flat plate (Fig. P15–58). Run for each of the available turbulence models. For each case, record CD. Is there much variation in the calculated values of CD? Which turbulence model yields the most correct value of drag coefficient? Discuss.
15–56 Repeat Prob. 15–55, except for fully developed turbulent flow through a smooth-walled pipe. Use FlowLab template Pipe_turbulent_developed. Calculate and plot the Darcy friction factor f as a function of Re. Compare f with that predicted in Chap. 8 for fully developed turbulent pipe flow through a smooth pipe. Discuss.
15–61 Consider laminar flow on a smooth heated flat plate (Fig. P15–61). Run FlowLab template Plate_laminar_temperature for two fluids: air and water. The inlet velocity is adjusted such that the Reynolds number for the air and water cases are approximately equal. Compare the 99 percent temperature thickness at the end of the plate to the 99 percent velocity thickness. Discuss your results. (Hint: What is the Prandtl number of air and of water?)
15–57 In Prob. 15–56, we considered fully developed turbulent flow through a smooth pipe. In this exercise, we examine fully developed turbulent flow through a rough pipe. Run FlowLab with template Pipe_turbulent_rough. Run several cases, each with a different value of normalized pipe roughness, e/D, but at the same Reynolds number. Calculate and tabulate Darcy friction factor f as a function of normalized roughness parameter e/D. Compare f with that predicted by the Colebrook equation for fully developed turbulent pipe flow in rough pipes. Discuss. 15–58 Consider the laminar boundary layer developing over a flat plate (Fig. P15–58). Run FlowLab with template Plate_laminar. The inlet velocity and length are chosen such that the Reynolds number at the end of the plate, ReL " rVL/m, is approximately 1 & 105, just on the verge of transition toward turbulence. From your CFD results, calculate the following, and compare to theory: (a) the boundary layer profile shape at x " L (compare to the Blasius profile), (b) boundary layer thickness d as a function of x, and (c) drag coefficient on the plate. Symmetry V Velocity inlet
Outflow outlet
V, T∞ Velocity inlet
Outflow outlet
Twall
x=0 Symmetry
x=L Wall
FIGURE P15–61 15–62 Repeat Prob. 15–61, except for turbulent flow on a smooth heated flat plate (Fig. P15–61). Use FlowLab template Plate_turbulent_temperature. Discuss the differences between the laminar and turbulent calculations. Specifically, which regime (laminar or turbulent) produces the largest variation between 99 percent temperature thickness and 99 percent velocity thickness? Explain. 15–63 Consider turbulent flow of water through a smooth, 90°, flanged elbow in a round pipe (Fig. P15–63). Because of symmetry, only half of the pipe is modeled; the center plane is specified as a “symmetry” boundary condition. The pipe walls are smooth. The inlet velocity and pipe diameter are chosen to yield a Reynolds number of 20,000. For the first D
x=0 Symmetry
Symmetry
x=L Wall
FIGURE P15–58 15–59 Repeat Prob. 15–58, but for turbulent flow on a smooth flat plate. Use FlowLab template Plate_turbulent. The Reynolds number at the end of the plate is approximately 1 & 107 for this case—well beyond the transition region.
B-B C-C
D-D
A-A Velocity inlet
FIGURE P15–63
E-E
F-F Pressure outlet
Section A-A:
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(default) case, the standard k-e turbulence model is used. Run FlowLab with template Elbow. This is a three-dimensional calculation, so expect significantly longer run times. The average pressure is calculated across several cross sections of the pipe: upstream of the elbow, in the elbow, and downstream of the elbow (sections A-A, B-B, etc., in Fig. P15–63). Plot average pressure as a function of axial distance along the pipe. Where does most of the pressure drop occur—in the pipe section upstream of the elbow, in the elbow itself, immediately downstream of the elbow, or in the pipe section downstream of the elbow? Discuss. 15–64 Run FlowLab with template Elbow, again using the standard k-e turbulence model. In this exercise, we study velocity vectors in the plane of several cross sections along the pipe. Compare the velocity vectors at a section upstream of the elbow, at a section in the elbow, and at several sections downstream of the elbow. At which locations do you observe counter-rotating eddies? How does the strength of the counter-rotating eddies change with downstream distance? Discuss. Explain why many manufacturers of pipe flowmeters recommend that their flowmeter be installed at least 10 or 20 pipe diameters downstream of an elbow. 15–65 Run FlowLab with template Elbow, again using the standard k-e turbulence model. In this exercise, we calculate the minor loss coefficient KL for the elbow of Prob. 15–63. In order to do so, we compare the pressure drop calculated through the pipe with the elbow to that through a straight pipe of the same overall length, and with identical inlet and outlet conditions. Calculate the pressure drop from inlet to outlet for both geometries. To calculate KL for the elbow, subtract )P of the straight pipe from )P of the pipe with the elbow. The difference thus represents the pressure drop due to the elbow alone. From this pressure drop and the average velocity through the pipe, calculate minor loss coefficient KL, and compare to the value given in Chap. 8 for a smooth, 90°, flanged elbow. 15–66 In this exercise, we examine the effect of the turbulence model on the calculation of the minor loss coefficient of a pipe elbow (Fig. P15–63). Using FlowLab template Elbow, repeat Prob. 15–65, but with various turbulence models. For each case, calculate KL. Is there much variation in the calculated values of KL? Which turbulence model yields the most correct value, compared with the empirical result of Chap. 8? The Spallart–Allmaras model is the simplest, while the Reynolds stress model is the most complicated of the four. Do the calculated results improve with turbulence model complexity? Discuss. 15–67 Consider flow over a two-dimensional airfoil of chord length Lc at an angle of attack a in a flow of freestream speed V with density r and viscosity m. Angle a is measured relative to the free-stream flow direction. (Fig. P15–67). In this exercise, we calculate the nondimensional lift and drag coefficients CL and CD that correspond to lift
and drag forces FL and FD, respectively. Free-stream velocity and chord length are chosen such that the Reynolds number based on V and Lc is 1 & 107 (turbulent boundary layer over nearly the entire airfoil). Run FlowLab with template Airfoil_angle at several values of a, ranging from #2 to 20°. For each case, calculate CL and CD. Plot CL and CD as functions of a. At approximately what angle of attack does this airfoil stall?
r, m
FL V
Lc
FD a
FIGURE P15–67 15–68 In this problem, we study the effect of Reynolds number on the lift and drag coefficients of an airfoil at various angles of attack. Note that the airfoil used here is of a different shape than that used in Problem 15–67. Run FlowLab with template “Airfoil_Reynolds.” For the case with Reynolds number equal to 3 & 106, calculate and plot CL and CD as functions of a, with a ranging from #2 to 24°. What is the stall angle for this case? Repeat for Re " 6 & 106. Compare the two results and discuss the effect of Reynolds number on lift and drag of this airfoil. 15–69 In this exercise, we examine the effect of grid resolution on the calculation of airfoil stall (flow separation) for the airfoil of Problem 15–67 at a " 15° and Re " 1 & 107. Run FlowLab with template Airfoil_mesh. Run for several levels of grid resolution. For each case, calculate CL and CD. How does grid resolution affect the stall angle? Has grid independence been achieved? 15–70 Consider creeping flow produced by the body of a microorganism swimming through water, represented here as
Far field inflow (velocity inlet)
V
Far field outflow (outflow)
Computational domain
Body surface (wall)
y
Axis
x
(axis)
R L
FIGURE P15–70
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a simple 2 & 1 ellipsoid (Fig. P15–70, not to scale). Applied boundary conditions are shown for each edge in parentheses. The flow is laminar, and the default values of V and L are chosen such that the Reynolds number Re " rVL/m is equal to 0.20. Run FlowLab with template Creep_domain. Vary the computational domain radius from R/L " 3 to 2000. For each case, calculate the drag coefficient CD on the body. How large of a computational domain is required for the drag coefficient to level off (far field boundary conditions no longer have significant influence)? Discuss. For the largest computational domain case (R/L " 2000), plot velocity vectors along a vertical line coincident with the y-axis. Compare to the velocity profile we would expect at very high Reynolds numbers. Discuss. 15–71 Run FlowLab with template Creep_Reynolds. In this exercise, the Reynolds number is varied from 0.1 to 100 for flow over an ellipsoid (Fig. P15–70). Plot CD as a function of Re, and compare velocity profiles along the y-axis as Re increases above the creeping flow regime. Discuss.
General CFD Problems* 15–72 Consider the two-dimensional wye of Fig. P15–72. Dimensions are in meters, and the drawing is not to scale. Incompressible flow enters from the left, and splits into two parts. Generate three coarse grids, with identical node distributions on all edges of the computational domain: (a) structured multiblock grid, (b) unstructured triangular grid, and (c) unstructured quadrilateral grid. Compare the number of cells in each case and comment about the quality of the grid in each case.
15–74 Repeat Prob. 15–73, except for turbulent flow of air with a uniform inlet velocity of 10.0 m/s. In addition, set the turbulence intensity at the inlet to 10 percent with a turbulent length scale of 0.5 m. Use the k-e turbulence model with wall functions. Set the outlet pressure at both outlets to the same value, and calculate the pressure drop through the wye. Also calculate the percentage of the inlet flow that goes out of each branch. Generate a plot of streamlines. Compare results with those of laminar flow (Prob. 15–73). 15–75 Generate a computational domain to study the laminar boundary layer growing on a flat plate at Re " 10,000. Generate a very coarse mesh, and then continually refine the mesh until the solution becomes grid independent. Discuss. 15–76 Repeat Prob. 15–75, except for a turbulent boundary layer at Re " 106. Discuss. 15–77 Generate a computational domain to study ventilation in a room (Fig. P15–77). Specifically, generate a rectangular room with a velocity inlet in the ceiling to model the supply air, and a pressure outlet in the ceiling to model the return air. You may make a two-dimensional approximation for simplicity (the room is infinitely long in the direction normal to the page in Fig. P15–77). Use a structured rectangular grid. Plot streamlines and velocity vectors. Discuss.
Air supply
Air return
(4.5, 3.5) (5, 3) (0, 1)
(2, 1)
(5, 0.5) (2.5, 0.5)
(0, 0)
(5, 0)
FIGURE P15–72 15–73 Choose one of the grids generated in Prob. 15–72, and run a CFD solution for laminar flow of air with a uniform inlet velocity of 0.02 m/s. Set the outlet pressure at both outlets to the same value, and calculate the pressure drop through the wye. Also calculate the percentage of the inlet flow that goes out of each branch. Generate a plot of streamlines.
* These problems require CFD software, although not any particular brand. Unlike the FlowLab problems of the previous section, these problems do not have premade templates. Instead, students must do the following problems “from scratch.”
FIGURE P15–77 15–78 Repeat Prob. 15–77, except use an unstructured triangular grid, keeping everything else the same. Do you get the same results as those of Prob. 15–77? Compare and discuss. 15–79 Repeat Prob. 15–77, except move the supply and/or return vents to various locations in the ceiling. Compare and discuss. 15–80 Choose one of the room geometries of Probs. 15–77 and 15–79, and add the energy equation to the calculations. In particular, model a room with air-conditioning, by specifying the supply air as cool (T " 18°C), while the walls, floor, and ceiling are warm (T " 26°C). Adjust the supply air speed until the average temperature in the room is as close as possible to 22°C. How much ventilation (in terms of number of room air volume changes per hour) is required to cool this room to an average temperature of 22°C? Discuss.
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15–81 Repeat Prob. 15–80, except create a three-dimensional room, with an air supply and an air return in the ceiling. Compare the two-dimensional results of Prob. 15–80 with the more realistic three-dimensional results of this problem. Discuss. 15–82 Generate a computational domain to study compressible flow of air through a converging nozzle with atmospheric pressure at the nozzle exit (Fig. P15–82). The nozzle walls may be approximated as inviscid (zero shear stress). Run several cases with various values of inlet pressure. How much inlet pressure is required to choke the flow? What happens if the inlet pressure is higher than this value? Discuss.
15–86 Repeat Prob. 15–85, except for turbulent, rather than laminar, flow. Compare to the laminar case. Which has the lower drag coefficient? Discuss. 15–87 Generate a computational domain to study Mach waves in a two-dimensional supersonic channel (Fig. P15–87). Specifically, the domain should consist of a simple rectangular channel with a supersonic inlet (Ma " 2.0), and with a very small bump on the lower wall. Using air with the inviscid flow approximation, generate a Mach wave, as sketched. Measure the Mach angle, and compare with theory (Chap. 12). Also discuss what happens when the Mach wave hits the opposite wall. Does it disappear, or does it reflect, and if so, what is the reflection angle? Discuss. Ma
Pressure inlet
?
Pressure outlet Bump
FIGURE P15–87 FIGURE P15–82 15–83 Repeat Prob. 15–82, except remove the inviscid flow approximation. Instead, let the flow be turbulent, with smooth, no-slip walls. Compare your results to those of Prob. 15–82. What is the major effect of friction in this problem? Discuss. 15–84 Generate a computational domain to study incompressible, laminar flow over a two-dimensional streamlined body (Fig. P15–84). Generate various body shapes, and calculate the drag coefficient for each shape. What is the smallest value of CD that you can achieve? (Note: For fun, this problem can be turned into a contest between students. Who can generate the lowest-drag body shape?)
V Body FD
FIGURE P15–84 15–85 Repeat Prob. 15–84, except for an axisymmetric, rather than a two-dimensional, body. Compare to the twodimensional case. Which has the lower drag coefficient? Discuss.
15–88 Repeat Prob. 15–87, except for several values of the Mach number, ranging from 1.10 to 3.0. Plot the calculated Mach angle as a function of Mach number and compare to the theoretical Mach angle (Chap. 12). Discuss.
Review Problems 15–89C For each statement, choose whether the statement is true or false, and discuss your answer briefly. (a) The physical validity of a CFD solution always improves as the grid is refined. (b) The x-component of the Navier–Stokes equation is an example of a transport equation. (c) For the same number of nodes in a two-dimensional mesh, a structured grid typically has fewer cells than an unstructured triangular grid. (d) A time-averaged turbulent flow CFD solution is only as good as the turbulence model used in the calculations. 15–90C In Prob. 15–18 we take advantage of top–bottom symmetry when constructing our computational domain and grid. Why can’t we also take advantage of the right–left symmetry in this exercise? Repeat the discussion for the case of potential flow. 15–91C Gerry creates the computational domain sketched in Fig. P15–91C to simulate flow through a sudden contraction in a two-dimensional duct. He is interested in the timeaveraged pressure drop (minor loss coefficient) created by the sudden contraction. Gerry generates a grid and calculates the flow with a CFD code, assuming steady, turbulent, incompressible flow (with a turbulence model).
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(a) Discuss one way that Gerry could improve his computational domain and grid so that he would get the same results in approximately half the computer time. (b) There may be a fundamental flaw in how Gerry has set up his computational domain. What is it? Discuss what should be different about Gerry’s setup.
In
Out
to 90° (vertical). Use identical inlet conditions and wall conditions for each case. Note that the second stage of heating elements should always be set to an angle of attack that is the negative of that of the first stage. Which angle of attack provides the most heat transfer to the air? Specifically, which angle of attack yields the highest average outlet temperature? Is this the same angle as calculated for the single-stage heat exchanger of Prob. 15–96? Discuss. 15–99 Generate a computational domain and grid, and calculate stationary turbulent flow over a spinning circular cylinder (Fig. P15–99). In which direction is the side force on the body—up or down? Explain. Plot streamlines in the flow. Where is the upstream stagnation point?
FIGURE P15–91C v
15–92C Think about modern high-speed, large-memory computer systems. What feature of such computers lends itself nicely to the solution of CFD problems using a multiblock grid with approximately equal numbers of cells in each individual block? Discuss. 15–93C What is the difference between multigridding and multiblocking? Discuss how each may be used to speed up a CFD calculation. Can these two be applied together? 15–94C Suppose you have a fairly complex geometry and a CFD code that can handle unstructured grids with triangular cells. Your grid generation code can create an unstructured grid very quickly. Give some reasons why it might be wiser to take the time to create a multiblock structured grid instead. In other words, is it worth the effort? Discuss. 15–95 Generate a computational domain and grid, and calculate flow through the single-stage heat exchanger of Prob. 15–21, with the heating elements set at a 45° angle of attack with respect to horizontal. Set the inlet air temperature to 20°C, and the wall temperature of the heating elements to 120°C. Calculate the average air temperature at the outlet. 15–96 Repeat the calculations of Prob. 15–95 for several angles of attack of the heating elements, from 0 (horizontal) to 90° (vertical). Use identical inlet conditions and wall conditions for each case. Which angle of attack provides the most heat transfer to the air? Specifically, which angle of attack yields the highest average outlet temperature? 15–97 Generate a computational domain and grid, and calculate flow through the two-stage heat exchanger of Prob. 15–24, with the heating elements of the first stage set at a 45° angle of attack with respect to horizontal, and those of the second stage set to an angle of attack of #45°. Set the inlet air temperature to 20°C, and the wall temperature of the heating elements to 120°C. Calculate the average air temperature at the outlet. 15–98 Repeat the calculations of Prob. 15–97 for several angles of attack of the heating elements, from 0 (horizontal)
V
D
FIGURE P15–99 15–100 For the spinning cylinder of Fig. P15–99, generate a dimensionless parameter for rotational speed relative to free-stream speed (combine variables v, D, and V into a nondimensional Pi group). Repeat the calculations of Prob. 15–99 for several values of angular velocity v. Use identical inlet conditions for each case. Plot lift and drag coefficients as functions of your dimensionless parameter. Discuss. 15–101 Consider the flow of air into a two-dimensional slot along the floor of a large room, where the floor is coincident with the x-axis (Fig. P15–101). Generate an appropriate computational domain and grid. Using the inviscid flow approximation, calculate vertical velocity component y as a function of distance away from the slot along the y-axis. Compare with the potential flow results of Chap. 10 for flow into a line sink. Discuss. y Room
Floor
⋅ V
x
FIGURE P15–101 15–102 For the slot flow of Prob. 15–101, change to laminar flow instead of inviscid flow, and recompute the flow field. Compare your results to the inviscid flow case and to the potential flow case of Chap. 10. Plot contours of vorticity.
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Where is the irrotational flow approximation appropriate? Discuss. 15–103 Generate a computational domain and grid, and calculate the flow of air into a two-dimensional vacuum cleaner inlet (Fig. P15–103), using the inviscid flow approximation. Compare your results with those predicted in Chap. 10 for potential flow. Discuss. 15–104 For the vacuum cleaner of Prob. 15–103, change to laminar flow instead of inviscid flow, and recompute the flow field. Compare your results to the inviscid flow case and to the potential flow case of Chap. 10. Discuss.
⋅
y
V Vacuum nozzle
w
b b
b x
Maximum speed, potential flow
FIGURE P15–103
Stagnation point
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APPENDIX
P R O P E R T Y TA B L E S A N D CHARTS (SI UNITS)*
1
TABLE A–1
Molar Mass, Gas Constant, and Ideal-Gas Specfic Heats of Some Substances 886 TABLE A–2 Boiling and Freezing Point Properties 887 TABLE A–3 Properties of Saturated Water 888 TABLE A–4 Properties of Saturated Refrigerant-134a 889 TABLE A–5 Properties of Saturated Ammonia 890 TABLE A–6 Properties of Saturated Propane 891 TABLE A–7 Properties of Liquids 892 TABLE A–8 Properties of Liquid Metals 893 TABLE A–9 Properties of Air at 1 atm Pressure 894 TABLE A–10 Properties of Gases at 1 atm Pressure 895 TABLE A–11 Properties of the Atmosphere at High Altitude 897 FIGURE A–12 The Moody Chart for the Friction Factor for Fully Developed Flow in Circular Pipes 898 TABLE A–13 One-dimensional isentropic compressible flow functions for an ideal gas with k ! 1.4 899 TABLE A–14 One-dimensional normal shock functions for an ideal gas with k ! 1.4 900 TABLE A–15 Rayleigh flow functions for an ideal gas with k ! 1.4 901 TABLE A–16 Fanno flow functions for an ideal gas with k ! 1.4 902
* Most properties in the tables are obtained from the property database of EES, and the original sources are listed under the tables. Properties are often listed to more significant digits than the claimed accuracy for the purpose of minimizing accumulated round-off error in hand calculations and ensuring a close match with the results obtained with EES.
885
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886 FLUID MECHANICS
TA B L E A – 1 Molar mass, gas constant, and ideal-gas specfic heats of some substances
Substance Air Ammonia, NH3 Argon, Ar Bromine, Br2 Isobutane, C4H10 n-Butane, C4H10 Carbon dioxide, CO2 Carbon monoxide, CO Chlorine, Cl2 Chlorodifluoromethane (R-22), CHClF2 Ethane, C2H6 Ethylene, C2H4 Fluorine, F2 Helium, He n-Heptane, C7H16 n-Hexane, C6H14 Hydrogen, H2 Krypton, Kr Methane, CH4 Neon, Ne Nitrogen, N2 Nitric oxide, NO Nitrogen dioxide, NO2 Oxygen, O2 n-Pentane, C5H12 Propane, C3H8 Propylene, C3H6 Steam, H2O Sulfur dioxide, SO2 Tetrachloromethane, CCl4 Tetrafluoroethane (R-134a), C2H2F4 Trifluoroethane (R-143a), C2H3F3 Xenon, Xe
Specific Heat Data at 25°C
Molar Mass M, kg/kmol
Gas Constant R, kJ/kg · K*
cp, kJ/kg · K
28.97 17.03 39.95 159.81 58.12 58.12 44.01 28.01 70.905 86.47 30.070 28.054 38.00 4.003 100.20 86.18 2.016 83.80 16.04 20.183 28.01 30.006 46.006 32.00 72.15 44.097 42.08 18.015 64.06 153.82 102.03 84.04 131.30
0.2870 0.4882 0.2081 0.05202 0.1430 0.1430 0.1889 0.2968 0.1173 0.09615 0.2765 0.2964 0.2187 2.077 0.08297 0.09647 4.124 0.09921 0.5182 0.4119 0.2968 0.2771 0.1889 0.2598 0.1152 0.1885 0.1976 0.4615 0.1298 0.05405 0.08149 0.09893 0.06332
1.005 2.093 0.5203 0.2253 1.663 1.694 0.8439 1.039 0.4781 0.6496 1.744 1.527 0.8237 5.193 1.649 1.654 14.30 0.2480 2.226 1.030 1.040 0.9992 0.8060 0.9180 1.664 1.669 1.531 1.865 0.6228 0.5415 0.8334 0.9291 0.1583
cv, kJ/kg · K 0.7180 1.605 0.3122 0.1732 1.520 1.551 0.6550 0.7417 0.3608 0.5535 1.468 1.231 0.6050 3.116 1.566 1.558 10.18 0.1488 1.708 0.6180 0.7429 0.7221 0.6171 0.6582 1.549 1.480 1.333 1.403 0.4930 0.4875 0.7519 0.8302 0.09499
k ! cp /cv 1.400 1.304 1.667 1.300 1.094 1.092 1.288 1.400 1.325 1.174 1.188 1.241 1.362 1.667 1.053 1.062 1.405 1.667 1.303 1.667 1.400 1.384 1.306 1.395 1.074 1.127 1.148 1.329 1.263 1.111 1.108 1.119 1.667
* The unit kJ/kg · K is equivalent to kPa · m3/kg · K. The gas constant is calculated from R ! Ru /M, where Ru ! 8.31447 kJ/kmol · K is the universal gas constant and M is the molar mass. Source: Specific heat values are obtained primarily from the property routines prepared by The National Institute of Standards and Technology (NIST), Gaithersburg, MD.
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887 APPENDIX 1
TA B L E A – 2 Boiling and freezing point properties Boiling Data at 1 atm
Substance
Normal Boiling Point, °C
Ammonia
"33.3
Latent Heat of Vaporization hfg , kJ/kg 1357
Freezing Point, °C "77.7
Latent Heat of Fusion hif , kJ/kg 322.4
Argon Benzene Brine (20% sodium chloride by mass) n-Butane Carbon dioxide Ethanol Ethyl alcohol Ethylene glycol Glycerine Helium Hydrogen Isobutane Kerosene Mercury Methane
103.9 "0.5 "78.4* 78.2 78.6 198.1 179.9 "268.9 "252.8 "11.7 204–293 356.7 "161.5
— 385.2 230.5 (at 0°C) 838.3 855 800.1 974 22.8 445.7 367.1 251 294.7 510.4
"17.4 "138.5 "56.6 "114.2 "156 "10.8 18.9 — "259.2 "160 "24.9 "38.9 "182.2
109 108 181.1 200.6 — 59.5 105.7 — 11.4 58.4
Methanol Nitrogen
64.5 "195.8
1100 198.6
"97.7 "210
99.2 25.3
124.8
306.3
"57.5
180.7
"218.8
13.7
"187.7
80.0
Octane Oil (light) Oxygen Petroleum Propane
"185.9 80.2
"183 — "42.1
Refrigerant-134a
"26.1
Water
100
161.6 394
Freezing Data
212.7 230–384 427.8 216.8
2257
"189.3 5.5
"96.6
0.0
28 126 — 80.3
—
333.7
Liquid Properties Temperature, °C
Density r, kg/m3
Specific Heat cp, kJ/kg · K
"33.3 "20 0 25 "185.6 20
682 665 639 602 1394 879
4.43 4.52 4.60 4.80 1.14 1.72
20 "0.5 0 25 20 20 20 "268.9 "252.8 "11.7 20 25 "161.5 "100 25 "195.8 "160 20 25 "183 20 "42.1 0 50 "50 "26.1 0 25 0 25 50 75 100
1150 601 298 783 789 1109 1261 146.2 70.7 593.8 820 13,560 423 301 787 809 596 703 910 1141 640 581 529 449 1443 1374 1295 1207 1000 997 988 975 958
3.11 2.31 0.59 2.46 2.84 2.84 2.32 22.8 10.0 2.28 2.00 0.139 3.49 5.79 2.55 2.06 2.97 2.10 1.80 1.71 2.0 2.25 2.53 3.13 1.23 1.27 1.34 1.43 4.22 4.18 4.18 4.19 4.22
* Sublimation temperature. (At pressures below the triple-point pressure of 518 kPa, carbon dioxide exists as a solid or gas. Also, the freezing-point temperature of carbon dioxide is the triple-point temperature of "56.5°C.)
cen72367_appx1.qxd 11/17/04 4:34 PM Page 888
888 FLUID MECHANICS
TA B L E A – 3 Properties of saturated water
Temp. T, °C
Saturation Pressure Psat, kPa
Density r, kg/m3
Enthalpy of Vaporization hfg, kJ/kg
Specific Heat cp, J/kg · K Liquid
Vapor
Thermal Conductivity k, W/m · K
Prandtl Number Pr
Dynamic Viscosity m, kg/m · s Vapor
Volume Expansion Coefficient b, 1/K Liquid
Liquid
Vapor
Liquid
Vapor
Liquid
Liquid
Vapor
0.01 5 10 15 20
0.6113 0.8721 1.2276 1.7051 2.339
999.8 999.9 999.7 999.1 998.0
0.0048 0.0068 0.0094 0.0128 0.0173
2501 2490 2478 2466 2454
4217 4205 4194 4186 4182
1854 1857 1862 1863 1867
0.561 0.571 0.580 0.589 0.598
0.0171 0.0173 0.0176 0.0179 0.0182
1.792 # 10"3 1.519 # 10"3 1.307 # 10"3 1.138 # 10"3 1.002 # 10"3
0.922 0.934 0.946 0.959 0.973
# # # # #
10"5 10"5 10"5 10"5 10"5
13.5 11.2 9.45 8.09 7.01
1.00 1.00 1.00 1.00 1.00
"0.068 0.015 0.733 0.138 0.195
# # # # #
10"3 10"3 10"3 10"3 10"3
25 30 35 40 45
3.169 4.246 5.628 7.384 9.593
997.0 996.0 994.0 992.1 990.1
0.0231 0.0304 0.0397 0.0512 0.0655
2442 2431 2419 2407 2395
4180 4178 4178 4179 4180
1870 1875 1880 1885 1892
0.607 0.615 0.623 0.631 0.637
0.0186 0.0189 0.0192 0.0196 0.0200
0.891 0.798 0.720 0.653 0.596
# # # # #
10"3 10"3 10"3 10"3 10"3
0.987 1.001 1.016 1.031 1.046
# # # # #
10"5 10"5 10"5 10"5 10"5
6.14 5.42 4.83 4.32 3.91
1.00 1.00 1.00 1.00 1.00
0.247 0.294 0.337 0.377 0.415
# # # # #
10"3 10"3 10"3 10"3 10"3
50 55 60 65 70
12.35 15.76 19.94 25.03 31.19
988.1 985.2 983.3 980.4 977.5
0.0831 0.1045 0.1304 0.1614 0.1983
2383 2371 2359 2346 2334
4181 4183 4185 4187 4190
1900 1908 1916 1926 1936
0.644 0.649 0.654 0.659 0.663
0.0204 0.0208 0.0212 0.0216 0.0221
0.547 0.504 0.467 0.433 0.404
# # # # #
10"3 10"3 10"3 10"3 10"3
1.062 1.077 1.093 1.110 1.126
# # # # #
10"5 10"5 10"5 10"5 10"5
3.55 3.25 2.99 2.75 2.55
1.00 1.00 1.00 1.00 1.00
0.451 0.484 0.517 0.548 0.578
# # # # #
10"3 10"3 10"3 10"3 10"3
75 80 85 90 95
38.58 47.39 57.83 70.14 84.55
974.7 971.8 968.1 965.3 961.5
0.2421 0.2935 0.3536 0.4235 0.5045
2321 2309 2296 2283 2270
4193 4197 4201 4206 4212
1948 1962 1977 1993 2010
0.667 0.670 0.673 0.675 0.677
0.0225 0.0230 0.0235 0.0240 0.0246
0.378 0.355 0.333 0.315 0.297
# # # # #
10"3 10"3 10"3 10"3 10"3
1.142 1.159 1.176 1.193 1.210
# # # # #
10"5 10"5 10"5 10"5 10"5
2.38 2.22 2.08 1.96 1.85
1.00 1.00 1.00 1.00 1.00
0.607 0.653 0.670 0.702 0.716
# # # # #
10"3 10"3 10"3 10"3 10"3
100 110 120 130 140
101.33 143.27 198.53 270.1 361.3
957.9 950.6 943.4 934.6 921.7
0.5978 0.8263 1.121 1.496 1.965
2257 2230 2203 2174 2145
4217 4229 4244 4263 4286
2029 2071 2120 2177 2244
0.679 0.682 0.683 0.684 0.683
0.0251 0.0262 0.0275 0.0288 0.0301
0.282 0.255 0.232 0.213 0.197
# # # # #
10"3 10"3 10"3 10"3 10"3
1.227 1.261 1.296 1.330 1.365
# # # # #
10"5 10"5 10"5 10"5 10"5
1.75 1.58 1.44 1.33 1.24
1.00 1.00 1.00 1.01 1.02
0.750 0.798 0.858 0.913 0.970
# # # # #
10"3 10"3 10"3 10"3 10"3
150 160 170 180 190
475.8 617.8 791.7 1,002.1 1,254.4
916.6 907.4 897.7 887.3 876.4
2.546 3.256 4.119 5.153 6.388
2114 2083 2050 2015 1979
4311 4340 4370 4410 4460
2314 2420 2490 2590 2710
0.682 0.680 0.677 0.673 0.669
0.0316 0.0331 0.0347 0.0364 0.0382
0.183 0.170 0.160 0.150 0.142
# # # # #
10"3 10"3 10"3 10"3 10"3
1.399 1.434 1.468 1.502 1.537
# # # # #
10"5 10"5 10"5 10"5 10"5
1.16 1.09 1.03 0.983 0.947
1.02 1.05 1.05 1.07 1.09
1.025 1.145 1.178 1.210 1.280
# # # # #
10"3 10"3 10"3 10"3 10"3
200 220 240 260 280
1,553.8 2,318 3,344 4,688 6,412
864.3 840.3 813.7 783.7 750.8
7.852 11.60 16.73 23.69 33.15
1941 1859 1767 1663 1544
4500 4610 4760 4970 5280
2840 3110 3520 4070 4835
0.663 0.650 0.632 0.609 0.581
0.0401 0.0442 0.0487 0.0540 0.0605
0.134 0.122 0.111 0.102 0.094
# # # # #
10"3 10"3 10"3 10"3 10"3
1.571 1.641 1.712 1.788 1.870
# # # # #
10"5 10"5 10"5 10"5 10"5
0.910 0.865 0.836 0.832 0.854
1.11 1.15 1.24 1.35 1.49
1.350 1.520 1.720 2.000 2.380
# # # # #
10"3 10"3 10"3 10"3 10"3
1405 1239 1028 720 0
5750 6540 8240 14,690 —
5980 7900 11,870 25,800 —
0.548 0.509 0.469 0.427 —
0.0695 0.0836 0.110 0.178 —
0.086 0.078 0.070 0.060 0.043
# # # # #
10"3 10"3 10"3 10"3 10"3
1.965 2.084 2.255 2.571 4.313
# # # # #
10"5 10"5 10"5 10"5 10"5
0.902 1.00 1.23 2.06
1.69 1.97 2.43 3.73
2.950 # 10"3
300 320 340 360 374.14
8,581 11,274 14,586 18,651 22,090
713.8 667.1 610.5 528.3 317.0
46.15 64.57 92.62 144.0 317.0
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The temperatures 0.01°C, 100°C, and 374.14°C are the triple-, boiling-, and critical-point temperatures of water, respectively. The properties listed above (except the vapor density) can be used at any pressure with negligible error except at temperatures near the critical-point value. Note 2: The unit kJ/kg · °C for specific heat is equivalent to kJ/kg · K, and the unit W/m · °C for thermal conductivity is equivalent to W/m · K. Source: Viscosity and thermal conductivity data are from J. V. Sengers and J. T. R. Watson, Journal of Physical and Chemical Reference Data 15 (1986), pp. 1291–1322. Other data are obtained from various sources or calculated.
cen72367_appx1.qxd 11/17/04 4:34 PM Page 889
889 APPENDIX 1
TA B L E A – 4 Properties of saturated refrigerant-134a
Temp. T, °C
Saturation Pressure P, kPa
Liquid
Vapor
Liquid
Vapor
Vapor
Volume Expansion Coefficient b, 1/K Liquid
"40 "35 "30 "25 "20
51.2 66.2 84.4 106.5 132.8
1418 1403 1389 1374 1359
2.773 3.524 4.429 5.509 6.787
225.9 222.7 219.5 216.3 213.0
1254 1264 1273 1283 1294
748.6 764.1 780.2 797.2 814.9
0.1101 0.1084 0.1066 0.1047 0.1028
0.00811 0.00862 0.00913 0.00963 0.01013
4.878 4.509 4.178 3.882 3.614
# # # # #
10"4 10"4 10"4 10"4 10"4
2.550 3.003 3.504 4.054 4.651
# # # # #
10"6 10"6 10"6 10"6 10"6
5.558 5.257 4.992 4.757 4.548
0.235 0.266 0.299 0.335 0.374
0.00205 0.00209 0.00215 0.00220 0.00227
0.01760 0.01682 0.01604 0.01527 0.01451
"15 "10 "5 0 5
164.0 200.7 243.5 293.0 349.9
1343 1327 1311 1295 1278
8.288 10.04 12.07 14.42 17.12
209.5 206.0 202.4 198.7 194.8
1306 1318 1330 1344 1358
833.5 853.1 873.8 895.6 918.7
0.1009 0.0989 0.0968 0.0947 0.0925
0.01063 0.01112 0.01161 0.01210 0.01259
3.371 3.150 2.947 2.761 2.589
# # # # #
10"4 10"4 10"4 10"4 10"4
5.295 5.982 6.709 7.471 8.264
# # # # #
10"6 10"6 10"6 10"6 10"6
4.363 4.198 4.051 3.919 3.802
0.415 0.459 0.505 0.553 0.603
0.00233 0.00241 0.00249 0.00258 0.00269
0.01376 0.01302 0.01229 0.01156 0.01084
10 15 20 25 30
414.9 488.7 572.1 665.8 770.6
1261 1244 1226 1207 1188
20.22 23.75 27.77 32.34 37.53
190.8 186.6 182.3 177.8 173.1
1374 1390 1408 1427 1448
943.2 969.4 997.6 1028 1061
0.0903 0.0880 0.0856 0.0833 0.0808
0.01308 0.01357 0.01406 0.01456 0.01507
2.430 2.281 2.142 2.012 1.888
# # # # #
10"4 10"4 10"4 10"4 10"4
9.081 9.915 1.075 1.160 1.244
# # # # #
10"6 10"6 10"5 10"5 10"5
3.697 3.604 3.521 3.448 3.383
0.655 0.708 0.763 0.819 0.877
0.00280 0.00293 0.00307 0.00324 0.00342
0.01014 0.00944 0.00876 0.00808 0.00742
35 40 45 50 55
887.5 1017.1 1160.5 1318.6 1492.3
1168 1147 1125 1102 1078
43.41 50.08 57.66 66.27 76.11
168.2 163.0 157.6 151.8 145.7
1471 1498 1529 1566 1608
1098 1138 1184 1237 1298
0.0783 0.0757 0.0731 0.0704 0.0676
0.01558 0.01610 0.01664 0.01720 0.01777
1.772 1.660 1.554 1.453 1.355
# # # # #
10"4 10"4 10"4 10"4 10"4
1.327 1.408 1.486 1.562 1.634
# # # # #
10"5 10"5 10"5 10"5 10"5
3.328 3.285 3.253 3.231 3.223
0.935 0.995 1.058 1.123 1.193
0.00364 0.00390 0.00420 0.00456 0.00500
0.00677 0.00613 0.00550 0.00489 0.00429
60 65 70 75 80
1682.8 1891.0 2118.2 2365.8 2635.2
1053 1026 996.2 964 928.2
87.38 100.4 115.6 133.6 155.3
139.1 132.1 124.4 115.9 106.4
1659 1722 1801 1907 2056
1372 1462 1577 1731 1948
0.0647 0.0618 0.0587 0.0555 0.0521
0.01838 0.01902 0.01972 0.02048 0.02133
1.260 1.167 1.077 9.891 9.011
# # # # #
10"4 10"4 10"4 10"5 10"5
1.704 1.771 1.839 1.908 1.982
# # # # #
10"5 10"5 10"5 10"5 10"5
3.229 3.255 3.307 3.400 3.558
1.272 1.362 1.471 1.612 1.810
0.00554 0.00624 0.00716 0.00843 0.01031
0.00372 0.00315 0.00261 0.00209 0.00160
85 90 95 100
2928.2 3246.9 3594.1 3975.1
887.1 837.7 772.5 651.7
95.4 82.2 64.9 33.9
2287 2701 3675 7959
2281 2865 4144 8785
0.0484 0.0444 0.0396 0.0322
0.02233 0.02357 0.02544 0.02989
8.124 7.203 6.190 4.765
# # # #
10"5 10"5 10"5 10"5
2.071 2.187 2.370 2.833
# # # #
10"5 10"5 10"5 10"5
3.837 4.385 5.746 11.77
2.116 2.658 3.862 8.326
0.01336 0.01911 0.03343 0.10047
0.00114 0.00071 0.00033 0.00004
Density r, kg/m3 Liquid
Vapor
182.3 217.8 269.3 376.3
Enthalpy of Vaporization hfg, kJ/kg
Specific Heat cp, J/kg · K
Thermal Conductivity k, W/m · K
Prandtl Number Pr
Dynamic Viscosity m, kg/m · s Liquid
Vapor
Liquid
Surface Tension, N/m
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit kJ/kg · °C for specific heat is equivalent to kJ/kg · K, and the unit W/m · °C for thermal conductivity is equivalent to W/m · K. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: R. Tillner-Roth and H. D. Baehr, “An International Standard Formulation for the Thermodynamic Properties of 1,1,1,2-Tetrafluoroethane (HFC-134a) for Temperatures from 170 K to 455 K and Pressures up to 70 MPa,” J. Phys. Chem, Ref. Data, Vol. 23, No. 5, 1994; M. J. Assael, N. K. Dalaouti, A. A. Griva, and J. H. Dymond, “Viscosity and Thermal Conductivity of Halogenated Methane and Ethane Refrigerants,” IJR, Vol. 22, pp. 525–535, 1999; NIST REFPROP 6 program (M. O. McLinden, S. A. Klein, E. W. Lemmon, and A. P. Peskin, Physical and Chemical Properties Division, National Institute of Standards and Technology, Boulder, CO 80303, 1995).
cen72367_appx1.qxd 11/17/04 4:34 PM Page 890
890 FLUID MECHANICS
TA B L E A – 5 Properties of saturated ammonia
Saturation Pressure P, kPa
Liquid
Liquid
Vapor
Liquid
Vapor
Volume Expansion Coefficient b, 1/K Liquid
"40 "30 "25 "20 "15
71.66 119.4 151.5 190.1 236.2
690.2 677.8 671.5 665.1 658.6
0.6435 1.037 1.296 1.603 1.966
1389 1360 1345 1329 1313
4414 4465 4489 4514 4538
2242 2322 2369 2420 2476
— — 0.5968 0.5853 0.5737
0.01792 0.01898 0.01957 0.02015 0.02075
2.926 2.630 2.492 2.361 2.236
# # # # #
10"4 10"4 10"4 10"4 10"4
7.957 8.311 8.490 8.669 8.851
# # # # #
10"6 10"6 10"6 10"6 10"6
— — 1.875 1.821 1.769
0.9955 1.017 1.028 1.041 1.056
0.00176 0.00185 0.00190 0.00194 0.00199
0.03565 0.03341 0.03229 0.03118 0.03007
"10 "5 0 5 10
290.8 354.9 429.6 516 615.3
652.1 645.4 638.6 631.7 624.6
2.391 2.886 3.458 4.116 4.870
1297 1280 1262 1244 1226
4564 4589 4617 4645 4676
2536 2601 2672 2749 2831
0.5621 0.5505 0.5390 0.5274 0.5158
0.02138 0.02203 0.02270 0.02341 0.02415
2.117 2.003 1.896 1.794 1.697
# # # # #
10"4 10"4 10"4 10"4 10"4
9.034 9.218 9.405 9.593 9.784
# # # # #
10"6 10"6 10"6 10"6 10"6
1.718 1.670 1.624 1.580 1.539
1.072 1.089 1.107 1.126 1.147
0.00205 0.00210 0.00216 0.00223 0.00230
0.02896 0.02786 0.02676 0.02566 0.02457
15 20 25 30 35
728.8 857.8 1003 1167 1351
617.5 610.2 602.8 595.2 587.4
5.729 6.705 7.809 9.055 10.46
1206 1186 1166 1144 1122
4709 4745 4784 4828 4877
2920 3016 3120 3232 3354
0.5042 0.4927 0.4811 0.4695 0.4579
0.02492 0.02573 0.02658 0.02748 0.02843
1.606 1.519 1.438 1.361 1.288
# # # # #
10"4 10"4 10"4 10"4 10"4
9.978 1.017 1.037 1.057 1.078
# # # # #
10"6 10"5 10"5 10"5 10"5
1.500 1.463 1.430 1.399 1.372
1.169 1.193 1.218 1.244 1.272
0.00237 0.00245 0.00254 0.00264 0.00275
0.02348 0.02240 0.02132 0.02024 0.01917
40 45 50 55 60
1555 1782 2033 2310 2614
579.4 571.3 562.9 554.2 545.2
12.03 13.8 15.78 18.00 20.48
1099 1075 1051 1025 997.4
4932 4993 5063 5143 5234
3486 3631 3790 3967 4163
0.4464 0.4348 0.4232 0.4116 0.4001
0.02943 0.03049 0.03162 0.03283 0.03412
1.219 1.155 1.094 1.037 9.846
# # # # #
10"4 10"4 10"4 10"4 10"5
1.099 1.121 1.143 1.166 1.189
# # # # #
10"5 10"5 10"5 10"5 10"5
1.347 1.327 1.310 1.297 1.288
1.303 1.335 1.371 1.409 1.452
0.00287 0.00301 0.00316 0.00334 0.00354
0.01810 0.01704 0.01598 0.01493 0.01389
65 70 75 80 85
2948 3312 3709 4141 4609
536.0 526.3 516.2 505.7 494.5
23.26 26.39 29.90 33.87 38.36
968.9 939.0 907.5 874.1 838.6
5340 5463 5608 5780 5988
4384 4634 4923 5260 5659
0.3885 0.3769 0.3653 0.3538 0.3422
0.03550 0.03700 0.03862 0.04038 0.04232
9.347 8.879 8.440 8.030 7.645
# # # # #
10"5 10"5 10"5 10"5 10"5
1.213 1.238 1.264 1.292 1.322
# # # # #
10"5 10"5 10"5 10"5 10"5
1.285 1.287 1.296 1.312 1.338
1.499 1.551 1.612 1.683 1.768
0.00377 0.00404 0.00436 0.00474 0.00521
0.01285 0.01181 0.01079 0.00977 0.00876
90 95 100
5116 5665 6257
482.8 470.2 456.6
43.48 49.35 56.15
800.6 759.8 715.5
6242 6561 6972
6142 6740 7503
0.3306 0.3190 0.3075
0.04447 0.04687 0.04958
7.284 # 10"5 6.946 # 10"5 6.628 # 10"5
1.354 # 10"5 1.389 # 10"5 1.429 # 10"5
1.375 1.429 1.503
1.871 1.999 2.163
0.00579 0.00652 0.00749
0.00776 0.00677 0.00579
Temp. T, °C
Density r, kg/m3 Vapor
Enthalpy of Vaporization hfg, kJ/kg
Specific Heat cp, J/kg · K Liquid
Vapor
Thermal Conductivity k, W/m · K
Prandtl Number Pr
Dynamic Viscosity m, kg/m · s Liquid
Vapor
Surface Tension, N/m
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit kJ/kg · °C for specific heat is equivalent to kJ/kg · K, and the unit W/m · °C for thermal conductivity is equivalent to W/m · K. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Tillner-Roth, Harms-Watzenberg, and Baehr, “Eine neue Fundamentalgleichung fur Ammoniak,” DKV-Tagungsbericht 20:167–181, 1993; Liley and Desai, “Thermophysical Properties of Refrigerants,” ASHRAE, 1993, ISBN 1-1883413-10-9.
cen72367_appx1.qxd 11/17/04 4:34 PM Page 891
891 APPENDIX 1
TA B L E A – 6 Properties of saturated propane
Temp. T, °C "120 "110 "100 "90 "80
Saturation Pressure P, kPa 0.4053 1.157 2.881 6.406 12.97
Liquid
Vapor
Liquid
Vapor
Vapor
Volume Expansion Coefficient b, 1/K Liquid
664.7 654.5 644.2 633.8 623.2
0.01408 0.03776 0.08872 0.1870 0.3602
498.3 489.3 480.4 471.5 462.4
2003 2021 2044 2070 2100
1115 1148 1183 1221 1263
0.1802 0.1738 0.1672 0.1606 0.1539
0.00589 0.00645 0.00705 0.00769 0.00836
6.136 5.054 4.252 3.635 3.149
# # # # #
10"4 10"4 10"4 10"4 10"4
4.372 4.625 4.881 5.143 5.409
# # # # #
10"6 10"6 10"6 10"6 10"6
6.820 5.878 5.195 4.686 4.297
0.827 0.822 0.819 0.817 0.817
0.00153 0.00157 0.00161 0.00166 0.00171
0.02630 0.02486 0.02344 0.02202 0.02062
0.6439 1.081 1.724 2.629 3.864
453.1 443.5 433.6 423.1 412.1
2134 2173 2217 2258 2310
1308 1358 1412 1471 1535
0.1472 0.1407 0.1343 0.1281 0.1221
0.00908 0.00985 0.01067 0.01155 0.01250
2.755 2.430 2.158 1.926 1.726
# # # # #
10"4 10"4 10"4 10"4 10"4
5.680 5.956 6.239 6.529 6.827
# # # # #
10"6 10"6 10"6 10"6 10"6
3.994 3.755 3.563 3.395 3.266
0.818 0.821 0.825 0.831 0.839
0.00177 0.00184 0.00192 0.00201 0.00213
0.01923 0.01785 0.01649 0.01515 0.01382
Density r, kg/m3 Liquid
Vapor
Enthalpy of Vaporization hfg, kJ/kg
Specific Heat cp, J/kg · K
Thermal Conductivity k, W/m · K
Prandtl Number Pr
Dynamic Viscosity m, kg/m · s Liquid
Vapor
Liquid
Surface Tension, N/m
"70 "60 "50 "40 "30
24.26 42.46 70.24 110.7 167.3
612.5 601.5 590.3 578.8 567.0
"20 "10 0 5 10
243.8 344.4 473.3 549.8 635.1
554.7 542.0 528.7 521.8 514.7
5.503 7.635 10.36 11.99 13.81
400.3 387.8 374.2 367.0 359.5
2368 2433 2507 2547 2590
1605 1682 1768 1814 1864
0.1163 0.1107 0.1054 0.1028 0.1002
0.01351 0.01459 0.01576 0.01637 0.01701
1.551 # 1.397 # 1.259 # 1.195 # 1.135 #
10"4 10"4 10"4 10"4 10"4
7.136 7.457 7.794 7.970 8.151
# # # # #
10"6 10"6 10"6 10"6 10"6
3.158 3.069 2.996 2.964 2.935
0.848 0.860 0.875 0.883 0.893
0.00226 0.00242 0.00262 0.00273 0.00286
0.01251 0.01122 0.00996 0.00934 0.00872
15 20 25 30 35
729.8 834.4 949.7 1076 1215
507.5 500.0 492.2 484.2 475.8
15.85 18.13 20.68 23.53 26.72
351.7 343.4 334.8 325.8 316.2
2637 2688 2742 2802 2869
1917 1974 2036 2104 2179
0.0977 0.0952 0.0928 0.0904 0.0881
0.01767 0.01836 0.01908 0.01982 0.02061
1.077 1.022 9.702 9.197 8.710
# # # # #
10"4 10"4 10"5 10"5 10"5
8.339 8.534 8.738 8.952 9.178
# # # # #
10"6 10"6 10"6 10"6 10"6
2.909 2.886 2.866 2.850 2.837
0.905 0.918 0.933 0.950 0.971
0.00301 0.00318 0.00337 0.00358 0.00384
0.00811 0.00751 0.00691 0.00633 0.00575
40 45 50 60 70
1366 1530 1708 2110 2580
467.1 458.0 448.5 427.5 403.2
30.29 34.29 38.79 49.66 64.02
306.1 295.3 283.9 258.4 228.0
2943 3026 3122 3283 3595
2264 2361 2473 2769 3241
0.0857 0.0834 0.0811 0.0765 0.0717
0.02142 0.02228 0.02319 0.02517 0.02746
8.240 7.785 7.343 6.487 5.649
# # # # #
10"5 10"5 10"5 10"5 10"5
9.417 9.674 9.950 1.058 1.138
# # # # #
10"6 10"6 10"6 10"5 10"5
2.828 2.824 2.826 2.784 2.834
0.995 1.025 1.061 1.164 1.343
0.00413 0.00448 0.00491 0.00609 0.00811
0.00518 0.00463 0.00408 0.00303 0.00204
80 90
3127 3769
373.0 329.1
84.28 118.6
189.7 133.2
4501 6977
4173 7239
0.0663 0.0595
0.03029 0.03441
4.790 # 10"5 3.807 # 10"5
1.249 # 10"5 1.448 # 10"5
3.251 4.465
1.722 3.047
0.01248 0.02847
0.00114 0.00037
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit kJ/kg · °C for specific heat is equivalent to kJ/kg · K, and the unit W/m · °C for thermal conductivity is equivalent to W/m · K. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Reiner Tillner-Roth, “Fundamental Equations of State,” Shaker, Verlag, Aachan, 1998; B. A. Younglove and J. F. Ely, “Thermophysical Properties of Fluids. II Methane, Ethane, Propane, Isobutane, and Normal Butane,” J. Phys. Chem. Ref. Data, Vol. 16, No. 4, 1987; G.R. Somayajulu, “A Generalized Equation for Surface Tension from the Triple-Point to the CriticalPoint,” International Journal of Thermophysics, Vol. 9, No. 4, 1988.
cen72367_appx1.qxd 11/17/04 4:34 PM Page 892
892 FLUID MECHANICS
TA B L E A – 7 Properties of liquids
Temp. T, °C
Density r, kg/m3
Specific Heat cp, J/kg · K
Thermal Conductivity k, W/m · K
Thermal Diffusivity a, m2/s
Dynamic Viscosity m, kg/m · s
Kinematic Viscosity n, m2/s
Prandtl Number Pr
Volume Expansion Coeff. b, 1/K
Methane (CH4) "160 "150 "140 "130 "120 "110 "100 "90
420.2 405.0 388.8 371.1 351.4 328.8 301.0 261.7
3492 3580 3700 3875 4146 4611 5578 8902
0.1863 0.1703 0.1550 0.1402 0.1258 0.1115 0.0967 0.0797
1.270 1.174 1.077 9.749 8.634 7.356 5.761 3.423
# # # # # # # #
10"7 10"7 10"7 10"8 10"8 10"8 10"8 10"8
1.133 9.169 7.551 6.288 5.257 4.377 3.577 2.761
10"4 10"5 10"5 10"5 10"5 10"5 10"5 10"5
2.699 2.264 1.942 1.694 1.496 1.331 1.188 1.055
# # # # # # # #
10"7 10"7 10"7 10"7 10"7 10"7 10"7 10"7
2.126 1.927 1.803 1.738 1.732 1.810 2.063 3.082
0.00352 0.00391 0.00444 0.00520 0.00637 0.00841 0.01282 0.02922
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
7.429 6.531 5.795 5.185 4.677 4.250
# # # # # #
10"7 10"7 10"7 10"7 10"7 10"7
7.414 6.622 5.980 5.453 5.018 4.655
0.00118 0.00120 0.00123 0.00127 0.00132 0.00137
# # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"5 10"5
1.360 8.531 5.942 4.420 3.432 2.743 2.233 1.836 1.509
# # # # # # # # #
10"6 10"7 10"7 10"7 10"7 10"7 10"7 10"7 10"7
12.65 8.167 6.079 4.963 4.304 3.880 3.582 3.363 3.256
0.00142 0.00150 0.00161 0.00177 0.00199 0.00232 0.00286 0.00385 0.00628
8.219 5.287 3.339 1.970 1.201 7.878 5.232 3.464 2.455
# # # # # # # # #
10"3 10"3 10"3 10"3 10"3 10"4 10"4 10"4 10"4
84,101 54,327 34,561 20,570 12,671 8,392 5,631 3,767 2,697
4.242 9.429 2.485 8.565 3.794 2.046 1.241 8.029 6.595
# # # # # # # # #
10"3 10"4 10"4 10"5 10"5 10"5 10"5 10"6 10"6
46,636 10,863 2,962 1,080 499.3 279.1 176.3 118.1 98.31
# # # # # # # #
Methanol [CH3(OH)] 20 30 40 50 60 70
788.4 779.1 769.6 760.1 750.4 740.4
2515 2577 2644 2718 2798 2885
0.1987 0.1980 0.1972 0.1965 0.1957 0.1950
1.002 9.862 9.690 9.509 9.320 9.128
# # # # # #
10"7 10"8 10"8 10"8 10"8 10"8
5.857 5.088 4.460 3.942 3.510 3.146
Isobutane (R600a) "100 "75 "50 "25 0 25 50 75 100
683.8 659.3 634.3 608.2 580.6 550.7 517.3 478.5 429.6
1881 1970 2069 2180 2306 2455 2640 2896 3361
0.1383 0.1357 0.1283 0.1181 0.1068 0.0956 0.0851 0.0757 0.0669
1.075 1.044 9.773 8.906 7.974 7.069 6.233 5.460 4.634
# # # # # # # # #
10"7 10"7 10"8 10"8 10"8 10"8 10"8 10"8 10"8
# # # # # # # # #
10"8
9.305 5.624 3.769 2.688 1.993 1.510 1.155 8.785 6.483
Glycerin 0 5 10 15 20 25 30 35 40
1276 1273 1270 1267 1264 1261 1258 1255 1252
2262 2288 2320 2354 2386 2416 2447 2478 2513
0.2820 0.2835 0.2846 0.2856 0.2860 0.2860 0.2860 0.2860 0.2863
9.773 9.732 9.662 9.576 9.484 9.388 9.291 9.195 9.101
10"8 10"8 10"8 10"8 10"8 10"8 10"8 10"8
10.49 6.730 4.241 2.496 1.519 0.9934 0.6582 0.4347 0.3073
Engine Oil (unused) 0 20 40 60 80 100 120 140 150
899.0 888.1 876.0 863.9 852.0 840.0 828.9 816.8 810.3
1797 1881 1964 2048 2132 2220 2308 2395 2441
0.1469 0.1450 0.1444 0.1404 0.1380 0.1367 0.1347 0.1330 0.1327
9.097 8.680 8.391 7.934 7.599 7.330 7.042 6.798 6.708
# # # # # # # # #
10"8 10"8 10"8 10"8 10"8 10"8 10"8 10"8 10"8
3.814 0.8374 0.2177 0.07399 0.03232 0.01718 0.01029 0.006558 0.005344
Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
0.00070 0.00070 0.00070 0.00070 0.00070 0.00070 0.00070 0.00070 0.00070
cen72367_appx1.qxd 11/17/04 4:34 PM Page 893
893 APPENDIX 1
TA B L E A – 8 Properties of liquid metals
Temp. T, °C
Density r, kg/m3
Specific Heat cp, J/kg · K
Thermal Conductivity k, W/m · K
Thermal Diffusivity a, m2/s
Dynamic Viscosity m, kg/m · s
Kinematic Viscosity n, m2/s
Prandtl Number Pr
Volume Expansion Coeff. b, 1/K
Mercury (Hg) Melting Point: "39°C 0 25 50 75 100 150 200 250 300
13595 13534 13473 13412 13351 13231 13112 12993 12873
140.4 139.4 138.6 137.8 137.1 136.1 135.5 135.3 135.3
8.18200 8.51533 8.83632 9.15632 9.46706 10.07780 10.65465 11.18150 11.68150
4.287 # 4.514 # 4.734 # 4.956 # 5.170 # 5.595 # 5.996 # 6.363 # 6.705 #
10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6
1.687 1.534 1.423 1.316 1.245 1.126 1.043 9.820 9.336
# # # # # # # # #
10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"4 10"4
1.241 1.133 1.056 9.819 9.326 8.514 7.959 7.558 7.252
# # # # # # # # #
10"7 10"7 10"7 10"8 10"8 10"8 10"8 10"8 10"8
0.0289 0.0251 0.0223 0.0198 0.0180 0.0152 0.0133 0.0119 0.0108
1.545 1.436 1.215 1.048 9.157
# # # # #
10"7 10"7 10"7 10"7 10"8
0.01381 0.01310 0.01154 0.01022 0.00906
2.167 1.976 1.814 1.702 1.589 1.475 1.360
# # # # # # #
10"7 10"7 10"7 10"7 10"7 10"7 10"7
0.02252 0.02048 0.01879 0.01771 0.01661 0.01549 0.01434
7.432 5.967 4.418 3.188 2.909 2.614
# # # # # #
10"7 10"7 10"7 10"7 10"7 10"7
0.01106 0.008987 0.006751 0.004953 0.004593 0.004202
4.213 3.456 2.652 2.304 2.126
# # # # #
10"7 10"7 10"7 10"7 10"7
0.006023 0.004906 0.00374 0.003309 0.003143
10"7 10"7 10"7 10"7 10"7 10"7
0.02102 0.01611 0.01161 0.00753 0.00665 0.00579
Bismuth (Bi) Melting Point: 271°C 350 400 500 600 700
9969 9908 9785 9663 9540
146.0 148.2 152.8 157.3 161.8
16.28 16.10 15.74 15.60 15.60
1.118 1.096 1.052 1.026 1.010
# # # # #
10"5 10"5 10"5 10"5 10"5
1.540 1.422 1.188 1.013 8.736
# # # # #
10"3 10"3 10"3 10"3 10"4
Lead (Pb) Melting Point: 327°C 400 450 500 550 600 650 700
10506 10449 10390 10329 10267 10206 10145
158 156 155 155 155 155 155
15.97 15.74 15.54 15.39 15.23 15.07 14.91
9.623 9.649 9.651 9.610 9.568 9.526 9.483
# # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"6
2.277 2.065 1.884 1.758 1.632 1.505 1.379
# # # # # # #
10"3 10"3 10"3 10"3 10"3 10"3 10"3
Sodium (Na) Melting Point: 98°C 100 200 300 400 500 600
927.3 902.5 877.8 853.0 828.5 804.0
1378 1349 1320 1296 1284 1272
85.84 80.84 75.84 71.20 67.41 63.63
6.718 6.639 6.544 6.437 6.335 6.220
# # # # # #
10"5 10"5 10"5 10"5 10"5 10"5
6.892 5.385 3.878 2.720 2.411 2.101
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
Potassium (K) Melting Point: 64°C 200 300 400 500 600
795.2 771.6 748.0 723.9 699.6
790.8 772.8 754.8 750.0 750.0
43.99 42.01 40.03 37.81 35.50
6.995 7.045 7.090 6.964 6.765
# # # # #
10"5 10"5 10"5 10"5 10"5
3.350 2.667 1.984 1.668 1.487
# # # # #
10"4 10"4 10"4 10"4 10"4
Sodium–Potassium (%22Na-%78K) Melting Point: "11°C 100 200 300 400 500 600
847.3 823.2 799.1 775.0 751.5 728.0
944.4 922.5 900.6 879.0 880.1 881.2
25.64 26.27 26.89 27.50 27.89 28.28
3.205 3.459 3.736 4.037 4.217 4.408
# # # # # #
10"5 10"5 10"5 10"5 10"5 10"5
5.707 4.587 3.467 2.357 2.108 1.859
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
6.736 5.572 4.339 3.041 2.805 2.553
# # # # # #
Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
1.810 1.810 1.810 1.810 1.810 1.810 1.815 1.829 1.854
# # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4
cen72367_appx1.qxd 11/17/04 4:34 PM Page 894
894 FLUID MECHANICS
TA B L E A – 9 Properties of air at 1 atm pressure Temp. T, °C
Density r, kg/m3
Specific Heat cp J/kg · K
Thermal Conductivity k, W/m · K
Thermal Diffusivity a, m2/s
Dynamic Viscosity m, kg/m · s
Kinematic Viscosity n, m2/s
Prandtl Number Pr
"150 "100 "50 "40 "30
2.866 2.038 1.582 1.514 1.451
983 966 999 1002 1004
0.01171 0.01582 0.01979 0.02057 0.02134
4.158 8.036 1.252 1.356 1.465
# # # # #
10"6 10"6 10"5 10"5 10"5
8.636 1.189 1.474 1.527 1.579
# # # # #
10"6 10"6 10"5 10"5 10"5
3.013 5.837 9.319 1.008 1.087
# # # # #
10"6 10"6 10"6 10"5 10"5
0.7246 0.7263 0.7440 0.7436 0.7425
"20 "10 0 5 10
1.394 1.341 1.292 1.269 1.246
1005 1006 1006 1006 1006
0.02211 0.02288 0.02364 0.02401 0.02439
1.578 # 1.696 # 1.818 # 1.880 # 1.944 #
10"5 10"5 10"5 10"5 10"5
1.630 1.680 1.729 1.754 1.778
# # # # #
10"5 10"5 10"5 10"5 10"5
1.169 1.252 1.338 1.382 1.426
# # # # #
10"5 10"5 10"5 10"5 10"5
0.7408 0.7387 0.7362 0.7350 0.7336
15 20 25 30 35
1.225 1.204 1.184 1.164 1.145
1007 1007 1007 1007 1007
0.02476 0.02514 0.02551 0.02588 0.02625
2.009 2.074 2.141 2.208 2.277
# # # # #
10"5 10"5 10"5 10"5 10"5
1.802 1.825 1.849 1.872 1.895
# # # # #
10"5 10"5 10"5 10"5 10"5
1.470 1.516 1.562 1.608 1.655
# # # # #
10"5 10"5 10"5 10"5 10"5
0.7323 0.7309 0.7296 0.7282 0.7268
40 45 50 60 70
1.127 1.109 1.092 1.059 1.028
1007 1007 1007 1007 1007
0.02662 0.02699 0.02735 0.02808 0.02881
2.346 2.416 2.487 2.632 2.780
# # # # #
10"5 10"5 10"5 10"5 10"5
1.918 1.941 1.963 2.008 2.052
# # # # #
10"5 10"5 10"5 10"5 10"5
1.702 1.750 1.798 1.896 1.995
# # # # #
10"5 10"5 10"5 10"5 10"5
0.7255 0.7241 0.7228 0.7202 0.7177
80 90 100 120 140
0.9994 0.9718 0.9458 0.8977 0.8542
1008 1008 1009 1011 1013
0.02953 0.03024 0.03095 0.03235 0.03374
2.931 3.086 3.243 3.565 3.898
# # # # #
10"5 10"5 10"5 10"5 10"5
2.096 2.139 2.181 2.264 2.345
# # # # #
10"5 10"5 10"5 10"5 10"5
2.097 2.201 2.306 2.522 2.745
# # # # #
10"5 10"5 10"5 10"5 10"5
0.7154 0.7132 0.7111 0.7073 0.7041
160 180 200 250 300
0.8148 0.7788 0.7459 0.6746 0.6158
1016 1019 1023 1033 1044
0.03511 0.03646 0.03779 0.04104 0.04418
4.241 4.593 4.954 5.890 6.871
# # # # #
10"5 10"5 10"5 10"5 10"5
2.420 2.504 2.577 2.760 2.934
# # # # #
10"5 10"5 10"5 10"5 10"5
2.975 3.212 3.455 4.091 4.765
# # # # #
10"5 10"5 10"5 10"5 10"5
0.7014 0.6992 0.6974 0.6946 0.6935
350 400 450 500 600
0.5664 0.5243 0.4880 0.4565 0.4042
1056 1069 1081 1093 1115
0.04721 0.05015 0.05298 0.05572 0.06093
7.892 8.951 1.004 1.117 1.352
# # # # #
10"5 10"5 10"4 10"4 10"4
3.101 3.261 3.415 3.563 3.846
# # # # #
10"5 10"5 10"5 10"5 10"5
5.475 6.219 6.997 7.806 9.515
# # # # #
10"5 10"5 10"5 10"5 10"5
0.6937 0.6948 0.6965 0.6986 0.7037
700 800 900 1000 1500 2000
0.3627 0.3289 0.3008 0.2772 0.1990 0.1553
1135 1153 1169 1184 1234 1264
0.06581 0.07037 0.07465 0.07868 0.09599 0.11113
1.598 1.855 2.122 2.398 3.908 5.664
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
4.111 4.362 4.600 4.826 5.817 6.630
# # # # # #
10"5 10"5 10"5 10"5 10"5 10"5
1.133 1.326 1.529 1.741 2.922 4.270
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
0.7092 0.7149 0.7206 0.7260 0.7478 0.7539
Note: For ideal gases, the properties cp, k, m, and Pr are independent of pressure. The properties r, n, and a at a pressure P (in atm) other than 1 atm are determined by multiplying the values of r at the given temperature by P and by dividing n and a by P. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Keenan, Chao, Keyes, Gas Tables, Wiley, 198; and Thermophysical Properties of Matter, Vol. 3: Thermal Conductivity, Y. S. Touloukian, P. E. Liley, S. C. Saxena, Vol. 11: Viscosity, Y. S. Touloukian, S. C. Saxena, and P. Hestermans, IFI/Plenun, NY, 1970, ISBN 0-306067020-8.
cen72367_appx1.qxd 11/17/04 4:34 PM Page 895
895 APPENDIX 1
TA B L E A – 1 0 Properties of gases at 1 atm pressure Temp. T, °C
Density r, kg/m3
Specific Heat cp J/kg · K
Thermal Conductivity k, W/m · K
"50 0 50 100 150 200 300 400 500 1000 1500 2000
2.4035 1.9635 1.6597 1.4373 1.2675 1.1336 0.9358 0.7968 0.6937 0.4213 0.3025 0.2359
746 811 866.6 914.8 957.4 995.2 1060 1112 1156 1292 1356 1387
0.01051 0.01456 0.01858 0.02257 0.02652 0.03044 0.03814 0.04565 0.05293 0.08491 0.10688 0.11522
Thermal Diffusivity a, m2/s
Dynamic Viscosity m, kg/m · s
Kinematic Viscosity n, m2/s
Prandtl Number Pr
Carbon Dioxide, CO2 5.860 9.141 1.291 1.716 2.186 2.698 3.847 5.151 6.600 1.560 2.606 3.521
# # # # # # # # # # # #
10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
1.129 1.375 1.612 1.841 2.063 2.276 2.682 3.061 3.416 4.898 6.106 7.322
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
4.699 7.003 9.714 1.281 1.627 2.008 2.866 3.842 4.924 1.162 2.019 3.103
# # # # # # # # # # # #
10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
0.8019 0.7661 0.7520 0.7464 0.7445 0.7442 0.7450 0.7458 0.7460 0.7455 0.7745 0.8815
1.378 1.629 1.863 2.080 2.283 2.472 2.812 3.111 3.379 4.557 6.321 9.826
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
9.012 1.303 1.764 2.274 2.830 3.426 4.722 6.136 7.653 1.700 3.284 6.543
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
0.7840 0.7499 0.7328 0.7239 0.7191 0.7164 0.7134 0.7111 0.7087 0.7080 0.7733 0.9302
8.564 1.028 1.191 1.345 1.491 1.630 1.886 2.119 2.334 3.281 4.434 6.360
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
9.774 1.436 1.969 2.567 3.227 3.944 5.529 7.297 9.228 2.136 4.022 7.395
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
0.8116 0.7494 0.7282 0.7247 0.7284 0.7344 0.7450 0.7501 0.7502 0.7331 0.7936 1.0386
7.293 8.391 9.427 1.041 1.136 1.228 1.403 1.570 1.730 2.455 3.099 3.690
# # # # # # # # # # # #
10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
6.624 9.329 1.240 1.582 1.957 2.365 3.274 4.302 5.443 1.272 2.237 3.414
# # # # # # # # # # # #
10"5 10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
0.6562 0.7071 0.7191 0.7196 0.7174 0.7155 0.7149 0.7179 0.7224 0.7345 0.7795 1.1717
Carbon Monoxide, CO "50 0 50 100 150 200 300 400 500 1000 1500 2000
1.5297 1.2497 1.0563 0.9148 0.8067 0.7214 0.5956 0.5071 0.4415 0.2681 0.1925 0.1502
1081 1048 1039 1041 1049 1060 1085 1111 1135 1226 1279 1309
0.01901 0.02278 0.02641 0.02992 0.03330 0.03656 0.04277 0.04860 0.05412 0.07894 0.10458 0.13833
1.149 1.739 2.407 3.142 3.936 4.782 6.619 8.628 1.079 2.401 4.246 7.034
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4
Methane, CH4 "50 0 50 100 150 200 300 400 500 1000 1500 2000
0.8761 0.7158 0.6050 0.5240 0.4620 0.4132 0.3411 0.2904 0.2529 0.1536 0.1103 0.0860
"50 0 50 100 150 200 300 400 500 1000 1500 2000
0.11010 0.08995 0.07603 0.06584 0.05806 0.05193 0.04287 0.03650 0.03178 0.01930 0.01386 0.01081
2243 2217 2302 2443 2611 2791 3158 3510 3836 5042 5701 6001
0.02367 0.03042 0.03766 0.04534 0.05344 0.06194 0.07996 0.09918 0.11933 0.22562 0.31857 0.36750
1.204 1.917 2.704 3.543 4.431 5.370 7.422 9.727 1.230 2.914 5.068 7.120
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4
Hydrogen, H2 12635 13920 14349 14473 14492 14482 14481 14540 14653 15577 16553 17400
0.1404 0.1652 0.1881 0.2095 0.2296 0.2486 0.2843 0.3180 0.3509 0.5206 0.6581 0.5480
1.009 1.319 1.724 2.199 2.729 3.306 4.580 5.992 7.535 1.732 2.869 2.914
# # # # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
(Continued)
cen72367_appx1.qxd 11/17/04 4:34 PM Page 896
896 FLUID MECHANICS
TA B L E A – 1 0 Properties of gases at 1 atm pressure (Continued) Temp. T, °C
Density r, kg/m3
Specific Heat cp J/kg · K
Thermal Conductivity k, W/m · K
Thermal Diffusivity a, m2/s
"50 0 50 100 150 200 300 400 500 1000 1500 2000
1.5299 1.2498 1.0564 0.9149 0.8068 0.7215 0.5956 0.5072 0.4416 0.2681 0.1925 0.1502
957.3 1035 1042 1041 1043 1050 1070 1095 1120 1213 1266 1297
0.02001 0.02384 0.02746 0.03090 0.03416 0.03727 0.04309 0.04848 0.05358 0.07938 0.11793 0.18590
1.366 1.843 2.494 3.244 4.058 4.921 6.758 8.727 1.083 2.440 4.839 9.543
"50 0 50 100 150 200 300 400 500 1000 1500 2000
1.7475 1.4277 1.2068 1.0451 0.9216 0.8242 0.6804 0.5793 0.5044 0.3063 0.2199 0.1716
984.4 928.7 921.7 931.8 947.6 964.7 997.1 1025 1048 1121 1165 1201
0.02067 0.02472 0.02867 0.03254 0.03637 0.04014 0.04751 0.05463 0.06148 0.09198 0.11901 0.14705
1.201 1.865 2.577 3.342 4.164 5.048 7.003 9.204 1.163 2.678 4.643 7.139
Dynamic Viscosity m, kg/m · s
Kinematic Viscosity n, m2/s
Prandtl Number Pr
Nitrogen, N2 # # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4
1.390 1.640 1.874 2.094 2.300 2.494 2.849 3.166 3.451 4.594 5.562 6.426
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
9.091 1.312 1.774 2.289 2.851 3.457 4.783 6.242 7.816 1.713 2.889 4.278
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
0.6655 0.7121 0.7114 0.7056 0.7025 0.7025 0.7078 0.7153 0.7215 0.7022 0.5969 0.4483
1.616 1.916 2.194 2.451 2.694 2.923 3.350 3.744 4.114 5.732 7.133 8.417
# # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
9.246 1.342 1.818 2.346 2.923 3.546 4.923 6.463 8.156 1.871 3.243 4.907
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4
0.7694 0.7198 0.7053 0.7019 0.7019 0.7025 0.7030 0.7023 0.7010 0.6986 0.6985 0.6873
7.187 8.956 1.078 1.265 1.456 1.650 2.045 2.446 2.847 4.762 6.411 7.808
# # # # # # # # # # # #
10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
7.305 1.114 1.587 2.150 2.806 3.556 5.340 7.498 1.002 2.761 5.177 8.084
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4
1.0047 1.0033 0.9944 0.9830 0.9712 0.9599 0.9401 0.9240 0.9108 0.8639 0.8233 0.7833
Oxygen, O2 # # # # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4
Water Vapor, H2O "50 0 50 100 150 200 300 400 500 1000 1500 2000
0.9839 0.8038 0.6794 0.5884 0.5189 0.4640 0.3831 0.3262 0.2840 0.1725 0.1238 0.0966
1892 1874 1874 1887 1908 1935 1997 2066 2137 2471 2736 2928
0.01353 0.01673 0.02032 0.02429 0.02861 0.03326 0.04345 0.05467 0.06677 0.13623 0.21301 0.29183
7.271 1.110 1.596 2.187 2.890 3.705 5.680 8.114 1.100 3.196 6.288 1.032
# # # # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"3
Note: For ideal gases, the properties cp, k, m, and Pr are independent of pressure. The properties r, n, and a at a pressure P (in atm) other than 1 atm are determined by multiplying the values of r at the given temperature by P and by dividing n and a by P. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
cen72367_appx1.qxd 11/17/04 4:34 PM Page 897
897 APPENDIX 1
TA B L E A – 1 1 Properties of the atmosphere at high altitude Altitude, m
Temperature, °C
Pressure, kPa
Gravity g, m/s2
Speed of Sound, m/s
Density, kg/m3
Viscosity m, kg/m · s
Thermal Conductivity, W/m · K
0 200 400 600 800
15.00 13.70 12.40 11.10 9.80
101.33 98.95 96.61 94.32 92.08
9.807 9.806 9.805 9.805 9.804
340.3 339.5 338.8 338.0 337.2
1.225 1.202 1.179 1.156 1.134
1.789 1.783 1.777 1.771 1.764
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0253 0.0252 0.0252 0.0251 0.0250
1000 1200 1400 1600 1800
8.50 7.20 5.90 4.60 3.30
89.88 87.72 85.60 83.53 81.49
9.804 9.803 9.802 9.802 9.801
336.4 335.7 334.9 334.1 333.3
1.112 1.090 1.069 1.048 1.027
1.758 1.752 1.745 1.739 1.732
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0249 0.0248 0.0247 0.0245 0.0244
2000 2200 2400 2600 2800
2.00 0.70 "0.59 "1.89 "3.19
79.50 77.55 75.63 73.76 71.92
9.800 9.800 9.799 9.799 9.798
332.5 331.7 331.0 330.2 329.4
1.007 0.987 0.967 0.947 0.928
1.726 1.720 1.713 1.707 1.700
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0243 0.0242 0.0241 0.0240 0.0239
3000 3200 3400 3600 3800
"4.49 "5.79 "7.09 "8.39 "9.69
70.12 68.36 66.63 64.94 63.28
9.797 9.797 9.796 9.796 9.795
328.6 327.8 327.0 326.2 325.4
0.909 0.891 0.872 0.854 0.837
1.694 1.687 1.681 1.674 1.668
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0238 0.0237 0.0236 0.0235 0.0234
4000 4200 4400 4600 4800
"10.98 "12.3 "13.6 "14.9 "16.2
61.66 60.07 58.52 57.00 55.51
9.794 9.794 9.793 9.793 9.792
324.6 323.8 323.0 322.2 321.4
0.819 0.802 0.785 0.769 0.752
1.661 1.655 1.648 1.642 1.635
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0233 0.0232 0.0231 0.0230 0.0229
5000 5200 5400 5600 5800
"17.5 "18.8 "20.1 "21.4 "22.7
54.05 52.62 51.23 49.86 48.52
9.791 9.791 9.790 9.789 9.785
320.5 319.7 318.9 318.1 317.3
0.736 0.721 0.705 0.690 0.675
1.628 1.622 1.615 1.608 1.602
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0228 0.0227 0.0226 0.0224 0.0223
6000 6200 6400 6600 6800
"24.0 "25.3 "26.6 "27.9 "29.2
47.22 45.94 44.69 43.47 42.27
9.788 9.788 9.787 9.786 9.785
316.5 315.6 314.8 314.0 313.1
0.660 0.646 0.631 0.617 0.604
1.595 1.588 1.582 1.575 1.568
# # # # #
10"5 10"5 10"5 10"5 10"5
0.0222 0.0221 0.0220 0.0219 0.0218
7000 8000 9000
"30.5 "36.9 "43.4
41.11 35.65 30.80
9.785 9.782 9.779
312.3 308.1 303.8
0.590 0.526 0.467
1.561 # 10"5 1.527 # 10"5 1.493 # 10"5
0.0217 0.0212 0.0206
10,000 12,000 14,000 16,000 18,000
"49.9 "56.5 "56.5 "56.5 "56.5
26.50 19.40 14.17 10.53 7.57
9.776 9.770 9.764 9.758 9.751
299.5 295.1 295.1 295.1 295.1
0.414 0.312 0.228 0.166 0.122
1.458 1.422 1.422 1.422 1.422
0.0201 0.0195 0.0195 0.0195 0.0195
# # # # #
10"5 10"5 10"5 10"5 10"5
Source: U.S. Standard Atmosphere Supplements, U.S. Government Printing Office, 1966. Based on year-round mean conditions at 45° latitude and varies with the time of the year and the weather patterns. The conditions at sea level (z ! 0) are taken to be P ! 101.325 kPa, T ! 15°C, r ! 1.2250 kg/m3, g ! 9.80665 m2/s.
0.008
0.009
0 0.003–0.03 0.0016 0.000033 0.000005 0.00085 0.0005 0.00015 0.000007 0.00015
0 0.9–9 0.5 0.01 0.0015 0.26 0.15 0.046 0.002 0.045
mm
Roughness, e
Smooth pipes e/D = 0
e/D = 0.000001
Fully rough turbulent flow ( f levels off )
e/D = 0.000005
0.00005
0.0001
0.0002
0.001 0.0008 0.0006 0.0004
0.002
0.004
0.01 0.008 0.006
0.015
0.02
0.03
0.05 0.04
L V2 . Friction factors in the The Moody chart for the friction factor for fully developed flow in circular pipes for use in the head loss relation h L ! f D 2g 1 e/D 2.51 ! "2 log10 a $ b. turbulent flow are evaluated from the Colebrook equation 3.7 Re 1f 1f
Reynolds number, Re
0.00001 3 4 5 6 7 103 2(10 ) 3 4 5 6 8 104 2(10 ) 3 4 5 6 8 105 2(10 ) 3 4 5 6 8 106 2(10 ) 3 4 5 6 8 107 2(10 ) 3 4 5 6 8 108
ft
Material
Glass, plastic Concrete Wood stave Rubber, smoothed Copper or brass tubing Cast iron Galvanized iron Wrought iron Stainless steel Commercial steel
e
0.01
64/R
0.015
0.02
f
0.025
inar
0.03
0.04
0.05
0.06
0.07
0.08
Laminar Transitional Turbulent flow flow flow
Lam f= low,
FIGURE A–12
Darcy friction factor, f
0.1 0.09
cen72367_appx1.qxd 11/17/04 4:34 PM Page 898
898 FLUID MECHANICS
Relative roughness, e/D
cen72367_appx1.qxd 11/17/04 4:34 PM Page 899
899 APPENDIX 1
k$1 B 2 $ (k " 1)Ma2 0.5(k$1)%(k"1) A 1 2 k"1 ! ca b a1 $ Ma2bd A* Ma k $ 1 2 "k%(k"1) P k"1 ! a1 $ Ma2b P0 2 "1%(k"1) r k"1 ! a1 $ Ma2b r0 2 "1 T k"1 ! a1 $ Ma2b T0 2
TA B L E A – 1 3
Ma* ! Ma
One-dimensional isentropic compressible flow functions for an ideal gas with k ! 1.4 Ma
Ma*
A/A*
P/P0
r/r0
T/T0
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0 5.0 '
0 0.1094 0.2182 0.3257 0.4313 0.5345 0.6348 0.7318 0.8251 0.9146 1.0000 1.1583 1.2999 1.4254 1.5360 1.6330 1.7179 1.7922 1.8571 1.9140 1.9640 2.2361 2.2495
& 5.8218 2.9635 2.0351 1.5901 1.3398 1.1882 1.0944 1.0382 1.0089 1.0000 1.0304 1.1149 1.2502 1.4390 1.6875 2.0050 2.4031 2.8960 3.5001 4.2346 25.000 '
1.0000 0.9930 0.9725 0.9395 0.8956 0.8430 0.7840 0.7209 0.6560 0.5913 0.5283 0.4124 0.3142 0.2353 0.1740 0.1278 0.0935 0.0684 0.0501 0.0368 0.0272 0.0019 0
1.0000 0.9950 0.9803 0.9564 0.9243 0.8852 0.8405 0.7916 0.7400 0.6870 0.6339 0.5311 0.4374 0.3557 0.2868 0.2300 0.1841 0.1472 0.1179 0.0946 0.0760 0.0113 0
1.0000 0.9980 0.9921 0.9823 0.9690 0.9524 0.9328 0.9107 0.8865 0.8606 0.8333 0.7764 0.7184 0.6614 0.6068 0.5556 0.5081 0.4647 0.4252 0.3894 0.3571 0.1667 0
3.0 A/A*
Compressible flow functions
2.5
2.0 Ma* 1.5
1.0 T/T0
0.5
r/r* P/P0 0 0
0.5
1.0
1.5 Ma
2.0
2.5
3.0
cen72367_appx1.qxd 11/17/04 4:34 PM Page 900
900 FLUID MECHANICS
T01 ! T02
TA B L E A – 1 4
(k " 1)Ma21 $ 2 Ma2 ! B 2kMa21 " k $ 1
One-dimensional normal shock functions for an ideal gas with k ! 1.4
P 2 1 $ kMa21 2kMa21 " k $ 1 ! ! P 1 1 $ kMa22 k$1 r 2 P 2%P 1 (k $ 1)Ma21 V1 ! ! ! r1 T2%T1 2 $ (k " 1)Ma21 V2 T2 2 $ Ma21(k " 1) ! T1 2 $ Ma22(k " 1) P 02 Ma1 1 $ Ma22(k " 1)%2 (k $ 1)/[2(k" 1)] ! c d P 01 Ma2 1 $ Ma21(k " 1)%2
P 02 (1 $ kMa21)[1 $ Ma22(k " 1)%2]k%(k"1) ! P1 1 $ kMa22
Ma1
Ma2
P2/P1
r2/r1
T2/T1
P02/P01
P02/P1
1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 4.0 5.0 &
1.0000 0.9118 0.8422 0.7860 0.7397 0.7011 0.6684 0.6405 0.6165 0.5956 0.5774 0.5613 0.5471 0.5344 0.5231 0.5130 0.5039 0.4956 0.4882 0.4814 0.4752 0.4350 0.4152 0.3780
1.0000 1.2450 1.5133 1.8050 2.1200 2.4583 2.8200 3.2050 3.6133 4.0450 4.5000 4.9783 5.4800 6.0050 6.5533 7.1250 7.7200 8.3383 8.9800 9.6450 10.3333 18.5000 29.000 &
1.0000 1.1691 1.3416 1.5157 1.6897 1.8621 2.0317 2.1977 2.3592 2.5157 2.6667 2.8119 2.9512 3.0845 3.2119 3.3333 3.4490 3.5590 3.6636 3.7629 3.8571 4.5714 5.0000 6.0000
1.0000 1.0649 1.1280 1.1909 1.2547 1.3202 1.3880 1.4583 1.5316 1.6079 1.6875 1.7705 1.8569 1.9468 2.0403 2.1375 2.2383 2.3429 2.4512 2.5632 2.6790 4.0469 5.8000 &
1.0000 0.9989 0.9928 0.9794 0.9582 0.9298 0.8952 0.8557 0.8127 0.7674 0.7209 0.6742 0.6281 0.5833 0.5401 0.4990 0.4601 0.4236 0.3895 0.3577 0.3283 0.1388 0.0617 0
1.8929 2.1328 2.4075 2.7136 3.0492 3.4133 3.8050 4.2238 4.6695 5.1418 5.6404 6.1654 6.7165 7.2937 7.8969 8.5261 9.1813 9.8624 10.5694 11.3022 12.0610 21.0681 32.6335 &
5.0 P02 /P1
P2 /P1
Normal shock functions
4.0
r2 /r1
3.0 T2/T1 2.0
1.0 Ma2 P02/P01
0 1.0
1.5
2.0 Ma1
2.5
3.0
cen72367_appx1.qxd 11/17/04 4:34 PM Page 901
901 APPENDIX 1
P0 2 $ (k " 1)Ma2 k%(k"1) k$1 a b ! 2 P*0 1 $ kMa k$1
Ma(1 $ k) 2 T !a b T* 1 $ kMa2
1$k P ! P* 1 $ kMa2
r* (1 $ k)Ma2 V ! ! r V* 1 $ kMa2
TA B L E A – 1 5 Rayleigh flow functions for an ideal gas with k ! 1.4 Ma
T0/T 0*
P0/P 0*
T/T*
P/P*
V/V*
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0
0.0000 0.0468 0.1736 0.3469 0.5290 0.6914 0.8189 0.9085 0.9639 0.9921 1.0000 0.9787 0.9343 0.8842 0.8363 0.7934 0.7561 0.7242 0.6970 0.6738 0.6540
1.2679 1.2591 1.2346 1.1985 1.1566 1.1141 1.0753 1.0431 1.0193 1.0049 1.0000 1.0194 1.0777 1.1756 1.3159 1.5031 1.7434 2.0451 2.4177 2.8731 3.4245
0.0000 0.0560 0.2066 0.4089 0.6151 0.7901 0.9167 0.9929 1.0255 1.0245 1.0000 0.9118 0.8054 0.7017 0.6089 0.5289 0.4611 0.4038 0.3556 0.3149 0.2803
2.4000 2.3669 2.2727 2.1314 1.9608 1.7778 1.5957 1.4235 1.2658 1.1246 1.0000 0.7958 0.6410 0.5236 0.4335 0.3636 0.3086 0.2648 0.2294 0.2004 0.1765
0.0000 0.0237 0.0909 0.1918 0.3137 0.4444 0.5745 0.6975 0.8101 0.9110 1.0000 1.1459 1.2564 1.3403 1.4046 1.4545 1.4938 1.5252 1.5505 1.5711 1.5882
3.5 P0 /P0* 3.0
2.5 Rayleigh flow functions
T0 (k $ 1)Ma2[2 $ (k " 1)Ma2] ! T 0* (1 $ kMa2)2
2.0 V/V* 1.5
1.0 T0 /T 0* 0.5
T/T * P/P*
0 0
0.5
1.0
1.5 Ma
2.0
2.5
3.0
cen72367_appx1.qxd 11/17/04 4:34 PM Page 902
902 FLUID MECHANICS
T0 ! T *0
TA B L E A – 1 6
r0 P0 1 2 $ (k " 1)Ma2 (k$1)/2(k"1) a b ! ! P* r* Ma k$1 0 0
Fanno flow functions for an ideal gas with k ! 1.4
k$1 T ! T* 2 $ (k " 1)Ma2
1/2 1 k$1 P ! a b 2 P* Ma 2 $ (k " 1)Ma
1/2 r* k$1 V ! Ma a ! b r V* 2 $ (k " 1)Ma2
fL* 1 " Ma2 k $ 1 (k $ 1)Ma2 ! $ ln D 2k kMa2 2 $ (k " 1)Ma2
Ma
P0/P 0*
T/T*
P/P*
V/V*
fL*/D
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0
& 5.8218 2.9635 2.0351 1.5901 1.3398 1.1882 1.0944 1.0382 1.0089 1.0000 1.0304 1.1149 1.2502 1.4390 1.6875 2.0050 2.4031 2.8960 3.5001 4.2346
1.2000 1.1976 1.1905 1.1788 1.1628 1.1429 1.1194 1.0929 1.0638 1.0327 1.0000 0.9317 0.8621 0.7937 0.7282 0.6667 0.6098 0.5576 0.5102 0.4673 0.4286
& 10.9435 5.4554 3.6191 2.6958 2.1381 1.7634 1.4935 1.2893 1.1291 1.0000 0.8044 0.6632 0.5568 0.4741 0.4082 0.3549 0.3111 0.2747 0.2441 0.2182
0.0000 0.1094 0.2182 0.3257 0.4313 0.5345 0.6348 0.7318 0.8251 0.9146 1.0000 1.1583 1.2999 1.4254 1.5360 1.6330 1.7179 1.7922 1.8571 1.9140 1.9640
& 66.9216 14.5333 5.2993 2.3085 1.0691 0.4908 0.2081 0.0723 0.0145 0.0000 0.0336 0.0997 0.1724 0.2419 0.3050 0.3609 0.4099 0.4526 0.4898 0.5222
3.0 P0 /P 0*
Fanno flow functions
2.5
2.0 V/V * 1.5
1.0 T/T * 0.5 fL*/D
P/P*
0 0
0.5
1.0
1.5 Ma
2.0
2.5
3.0
cen72367_appx2.qxd 11/17/04 4:34 PM Page 903
APPENDIX
P R O P E R T Y TA B L E S AND CHARTS (ENGLISH UNITS)*
2
TABLE A–1E
Molar Mass, Gas Constant, and Ideal-Gas Specific Heats of Some Substances 904 TABLE A–2E Boiling and Freezing Point Properties 905 TABLE A–3E Properties of Saturated Water 906 TABLE A–4E Properties of Saturated Refrigerant-134a 907 TABLE A–5E Properties of Saturated Ammonia 908 TABLE A–6E Properties of Saturated Propane 909 TABLE A–7E Properties of Liquids 910 TABLE A–8E Properties of Liquid Metals 911 TABLE A–9E Properties of Air at 1 atm Pressure 912 TABLE A–10E Properties of Gases at 1 atm Pressure 913 TABLE A–11E Properties of the Atmosphere at High Altitude 915
*Most properties in the tables are obtained from the property database of EES, and the original sources are listed under the tables. Properties are often listed to more significant digits than the claimed accuracy for the purpose of minimizing accumulated round-off error in hand calculations and ensuring a close match with the results obtained with EES.
903
cen72367_appx2.qxd 11/17/04 4:34 PM Page 904
904 FLUID MECHANICS
TA B L E A – 1 E Molar mass, gas constant, and ideal-gas specific heats of some substances Gas Constant R* Substance Air Ammonia, NH3 Argon, Ar Bromine, Br2 Isobutane, C4H10 n-Butane, C4H10 Carbon dioxide, CO2 Carbon monoxide, CO Chlorine, Cl2 Chlorodifluoromethane (R-22), CHClF2 Ethane, C2H6 Ethylene, C2H4 Fluorine, F2 Helium, He n-Heptane, C7H16 n-Hexane, C6H14 Hydrogen, H2 Krypton, Kr Methane, CH4 Neon, Ne Nitrogen, N2 Nitric oxide, NO Nitrogen dioxide, NO2 Oxygen, O2 n-Pentane, C5H12 Propane, C3H8 Propylene, C3H6 Steam, H2O Sulfur dioxide, SO2 Tetrachloromethane, CCl4 Tetrafluoroethane (R-134a), C2H2F4 Trifluoroethane (R-143a), C2H3F3 Xenon, Xe
ft3/
Specific Heat Data at 77°F
Molar Mass, M, lbm/lbmol
Btu/ lbm · R
psia · lbm · R
cp, Btu/lbm · R
cv , Btu/lbm · R
k ! cp /cv
28.97 17.03 39.95 159.81 58.12 58.12 44.01 28.01 70.905 86.47 30.070 28.054 38.00 4.003 100.20 86.18 2.016 83.80 16.04 20.183 28.01 30.006 46.006 32.00 72.15 44.097 42.08 18.015 64.06 153.82 102.03 84.04 131.30
0.06855 0.1166 0.04970 0.01242 0.03415 0.03415 0.04512 0.07089 0.02802 0.02297 0.06604 0.07079 0.05224 0.4961 0.01982 0.02304 0.9850 0.02370 0.1238 0.09838 0.07089 0.06618 0.04512 0.06205 0.02752 0.04502 0.04720 0.1102 0.03100 0.01291 0.01946 0.02363 0.01512
0.3704 0.6301 0.2686 0.06714 0.1846 0.1846 0.2438 0.3831 0.1514 0.1241 0.3569 0.3826 0.2823 2.681 0.1071 0.1245 5.323 0.1281 0.6688 0.5316 0.3831 0.3577 0.2438 0.3353 0.1487 0.2433 0.2550 0.5957 0.1675 0.06976 0.1052 0.1277 0.08173
0.2400 0.4999 0.1243 0.0538 0.3972 0.4046 0.2016 0.2482 0.1142 0.1552 0.4166 0.3647 0.1967 1.2403 0.3939 0.3951 3.416 0.05923 0.5317 0.2460 0.2484 0.2387 0.1925 0.2193 0.3974 0.3986 0.3657 0.4455 0.1488 0.1293 0.1991 0.2219 0.03781
0.1715 0.3834 0.07457 0.04137 0.3631 0.3705 0.1564 0.1772 0.08618 0.1322 0.3506 0.2940 0.1445 0.7442 0.3740 0.3721 2.431 0.03554 0.4080 0.1476 0.1774 0.1725 0.1474 0.1572 0.3700 0.3535 0.3184 0.3351 0.1178 0.1164 0.1796 0.1983 0.02269
1.400 1.304 1.667 1.300 1.094 1.092 1.288 1.400 1.325 1.174 1.188 1.241 1.362 1.667 1.053 1.062 1.405 1.667 1.303 1.667 1.400 1.384 1.306 1.395 1.074 1.127 1.148 1.329 1.263 1.111 1.108 1.119 1.667
*The gas constant is calculated from R ! Ru /M, where Ru ! 1.9859 Btu/lbmol · R ! 10.732 psia · ft3/lbmol · R is the universal gas constant and M is the molar mass. Source: Specific heat values are mostly obtained from the property routines prepared by The National Institute of Standards and Technology (NIST), Gaithersburg, MD.
cen72367_appx2.qxd 11/17/04 4:34 PM Page 905
905 APPENDIX 2
TA B L E A – 2 E Boiling and freezing point properties Boiling Data at 1 atm
Substance
Normal Boiling Point, °F
Ammonia
"27.9
Freezing Data
Latent Heat of Vaporization hfg, Btu/lbm
Freezing Point, °F
Latent Heat of Fusion hif, Btu/lbm
24.54
"107.9
138.6
Argon Benzene Brine (20% sodium chloride by mass) n-Butane Carbon dioxide Ethanol Ethyl alcohol Ethylene glycol Glycerine Helium Hydrogen Isobutane Kerosene Mercury Methane
"302.6 176.4
69.5 169.4
"308.7 41.9
12.0 54.2
219.0 31.1 "109.2* 172.8 173.5 388.6 355.8 "452.1 "423.0 10.9 399–559 674.1 "258.7
— 165.6 99.6 (at 32°F) 360.5 368 344.0 419 9.80 191.7 157.8 108 126.7 219.6
0.7 "217.3 "69.8 "173.6 "248.8 12.6 66.0 — "434.5 "255.5 "12.8 "38.0 296.0
— 34.5 — 46.9 46.4 77.9 86.3 — 25.6 45.5 — 4.90 25.1
Methanol Nitrogen
148.1 "320.4
473 85.4
"143.9 "346.0
42.7 10.9
Octane Oil (light) Oxygen Petroleum Propane
256.6 — "297.3 — "43.7
131.7 — 91.5 99–165 184.0
"71.5
77.9
"361.8
5.9
"305.8
34.4
"141.9
—
32
143.5
Refrigerant-134a
"15.0
Water
212
93.2
970.5
Liquid Properties Temperature, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
"27.9 0 40 80 "302.6 68
42.6 41.3 39.5 37.5 87.0 54.9
1.06 1.083 1.103 1.135 0.272 0.411
68 31.1 32 77 68 68 68 "452.1 "423.0 10.9 68 77 "258.7 "160 77 "320.4 "260 68 77 "297.3 68 "43.7 32 100 "40 "15 32 90 32 90 150 212
71.8 37.5 57.8 48.9 49.3 69.2 78.7 9.13 4.41 37.1 51.2 847 26.4 20.0 49.1 50.5 38.2 43.9 56.8 71.2 40.0 36.3 33.0 29.4 88.5 86.0 80.9 73.6 62.4 62.1 61.2 59.8
0.743 0.552 0.583 0.588 0.678 0.678 0.554 5.45 2.39 0.545 0.478 0.033 0.834 1.074 0.609 0.492 0.643 0.502 0.430 0.408 0.478 0.538 0.604 0.673 0.283 0.294 0.318 0.348 1.01 1.00 1.00 1.01
*Sublimation temperature. (At pressures below the triple-point pressure of 75.1 psia, carbon dioxide exists as a solid or gas. Also, the freezing-point temperature of carbon dioxide is the triple-point temperature of "69.8°F.)
cen72367_appx2.qxd 11/17/04 4:34 PM Page 906
906 FLUID MECHANICS
TA B L E A – 3 E Properties of saturated water
Temp. T, °F 32.02 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 212 220 230 240 250 260 270 280 290 300 320 340 360 380 400 450 500 550 600 650 700 705.44
Saturation Pressure Psat, psia
Liquid
Vapor
Enthalpy of Vaporization hfg, Btu/lbm
0.0887 0.1217 0.1780 0.2563 0.3632 0.5073 0.6988 0.9503 1.2763 1.6945 2.225 2.892 3.722 4.745 5.996 7.515 9.343 11.53 14.125 14.698 17.19 20.78 24.97 29.82 35.42 41.85 49.18 57.53 66.98 89.60 117.93 152.92 195.60 241.1 422.1 680.0 1046.7 1541 2210 3090 3204
62.41 62.42 62.41 62.36 62.30 62.22 62.12 62.00 61.86 61.71 61.55 61.38 61.19 60.99 60.79 60.57 60.35 60.12 59.87 59.82 59.62 59.36 59.09 58.82 58.53 58.24 57.94 57.63 57.31 56.65 55.95 55.22 54.46 53.65 51.46 48.95 45.96 42.32 37.31 27.28 19.79
0.00030 0.00034 0.00059 0.00083 0.00115 0.00158 0.00214 0.00286 0.00377 0.00493 0.00636 0.00814 0.0103 0.0129 0.0161 0.0199 0.0244 0.0297 0.0359 0.0373 0.0432 0.0516 0.0612 0.0723 0.0850 0.0993 0.1156 0.3390 0.1545 0.2033 0.2637 0.3377 0.4275 0.5359 0.9082 1.479 4.268 3.736 6.152 13.44 19.79
1075 1071 1065 1060 1054 1048 1043 1037 1031 1026 1020 1014 1008 1002 996 990 984 978 972 970 965 959 952 946 939 932 925 918 910 895 880 863 845 827 775 715 641 550 422 168 0
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft ·R
Dynamic Viscosity m, lbm/ft · s
Liquid
Vapor
Liquid
Vapor
Liquid
Vapor
1.010 1.004 1.000 0.999 0.999 0.999 0.999 0.999 0.999 0.999 0.999 0.999 1.000 1.000 1.001 1.002 1.004 1.005 1.007 1.007 1.009 1.011 1.013 1.015 1.018 1.020 1.023 1.026 1.029 1.036 1.044 1.054 1.065 1.078 1.121 1.188 1.298 1.509 2.086 13.80 $
0.446 0.447 0.448 0.449 0.450 0.451 0.453 0.454 0.456 0.458 0.460 0.463 0.465 0.468 0.472 0.475 0.479 0.483 0.487 0.488 0.492 0.497 0.503 0.509 0.516 0.523 0.530 0.538 0.547 0.567 0.590 0.617 0.647 0.683 0.799 0.972 1.247 1.759 3.103 25.90 $
0.324 0.329 0.335 0.341 0.347 0.352 0.358 0.363 0.367 0.371 0.375 0.378 0.381 0.384 0.386 0.388 0.390 0.391 0.392 0.392 0.393 0.394 0.394 0.395 0.395 0.395 0.395 0.395 0.394 0.393 0.391 0.389 0.385 0.382 0.370 0.352 0.329 0.299 0.267 0.254 $
0.0099 0.0100 0.0102 0.0104 0.0106 0.0108 0.0110 0.0112 0.0115 0.0117 0.0120 0.0122 0.0125 0.0128 0.0131 0.0134 0.0137 0.0141 0.0144 0.0145 0.0148 0.0152 0.0156 0.0160 0.0164 0.0168 0.0172 0.0177 0.0182 0.0191 0.0202 0.0213 0.0224 0.0237 0.0271 0.0312 0.0368 0.0461 0.0677 0.1964 $
1.204 # 10"3 1.038 # 10"3 8.781 # 10"4 7.536 # 10"4 6.556 # 10"4 5.764 # 10"4 5.117 # 10"4 4.578 # 10"4 4.128 # 10"4 3.744 # 10"4 3.417 # 10"4 3.136 # 10"4 2.889 # 10"4 2.675 # 10"4 2.483 # 10"4 2.317 # 10"4 2.169 # 10"4 2.036 # 10"4 1.917 # 10"4 1.894 # 10"4 1.808 #10"4 1.711 # 10"4 1.625 # 10"4 1.544 # 10"4 1.472 # 10"4 1.406 # 10"4 1.344 # 10"4 1.289 # 10"4 1.236 # 10"4 1.144 # 10"4 1.063 # 10"4 9.972 # 10"5 9.361 # 10"5 8.833 # 10"5 7.722 # 10"5 6.833 # 10"5 6.083 # 10"5 5.389 # 10"5 4.639 # 10"5 3.417 # 10"5 2.897 # 10"5
6.194 # 10"6 6.278 # 10"6 6.361 # 10"6 6.444 # 10"6 6.556 # 10"6 6.667 # 10"6 6.778 # 10"6 6.889 # 10"6 7.000 # 10"6 7.111 # 10"6 7.222 # 10"6 7.333 # 10"6 7.472 # 10"6 7.583 # 10"6 7.722 # 10"6 7.833 # 10"6 7.972 #10"6 8.083 # 10"6 8.222 # 10"6 8.250 #10"6 8.333 # 10"6 8.472 # 10"6 8.611 # 10"6 8.611 # 10"6 8.861 # 10"6 9.000 # 10"6 9.111 # 10"6 9.250 # 10"6 9.389 # 10"6 9.639 # 10"6 9.889 # 10"6 1.013 # 10"5 1.041 # 10"5 1.066 # 10"5 1.130 # 10"5 1.200 # 10"5 1.280 # 10"5 1.380 # 10"5 1.542 # 10"5 2.044 # 10"5 2.897 # 10"5
Prandtl Number Pr Liquid
Vapor
13.5 11.4 9.44 7.95 6.79 5.89 5.14 4.54 4.05 3.63 3.28 2.98 2.73 2.51 2.90 2.15 2.01 1.88 1.77 1.75 1.67 1.58 1.50 1.43 1.37 1.31 1.25 1.21 1.16 1.09 1.02 0.973 0.932 0.893 0.842 0.830 0.864 0.979 1.30 6.68
1.00 1.01 1.01 1.00 1.00 1.00 1.00 1.01 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.01 1.01 1.01 1.02 1.03 1.04 1.06 1.08 1.11 1.20 1.35 1.56 1.90 2.54 9.71
Volume Expansion Coefficient b, 1/R Liquid "0.038 0.003 0.047 0.080 0.115 0.145 0.174 0.200 0.224 0.246 0.267 0.287 0.306 0.325 0.346 0.367 0.382 0.395 0.412 0.417 0.429 0.443 0.462 0.480 0.497 0.514 0.532 0.549 0.566 0.636 0.656 0.681 0.720 0.771 0.912 1.111 1.445 1.885
# # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # #
10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The temperatures 32.02°F, 212°F, and 705.44°F are the triple-, boiling-, and critical-point temperatures of water, respectively. All properties listed above (except the vapor density) can be used at any pressure with negligible error except at temperatures near the critical-point value. Note 2: The unit Btu/lbm · °F for specific heat is equivalent to Btu/lbm · R, and the unit Btu/h · ft · °F for thermal conductivity is equivalent to Btu/h · ft · R. Source: Viscosity and thermal conductivity data are from J. V. Sengers and J. T. R. Watson, Journal of Physical and Chemical Reference Data 15 (1986), pp. 1291–1322. Other data are obtained from various sources or calculated.
cen72367_appx2.qxd 11/17/04 4:34 PM Page 907
907 APPENDIX 2
TA B L E A – 4 E Properties of saturated refrigerant-134a
Temp. T, °F
Saturation Pressure P, psia
Liquid
Vapor
Enthalpy of Vaporization hfg, Btu/lbm
"40 "30 "20 "10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210
7.4 9.9 12.9 16.6 21.2 26.6 33.1 40.8 49.8 60.2 72.2 85.9 101.4 119.1 138.9 161.2 186.0 213.5 244.1 277.8 314.9 355.8 400.7 449.9 504.0 563.8
88.51 87.5 86.48 85.44 84.38 83.31 82.2 81.08 79.92 78.73 77.51 76.25 74.94 73.59 72.17 70.69 69.13 67.48 65.72 63.83 61.76 59.47 56.85 53.75 49.75 43.19
0.1731 0.2258 0.2905 0.3691 0.4635 0.5761 0.7094 0.866 1.049 1.262 1.509 1.794 2.122 2.5 2.935 3.435 4.012 4.679 5.455 6.367 7.45 8.762 10.4 12.53 15.57 21.18
97.1 95.6 94.1 92.5 90.9 89.3 87.5 85.8 83.9 82.0 80.0 78.0 75.8 73.5 71.1 68.5 65.8 62.9 59.8 56.4 52.7 48.5 43.7 38.0 30.7 18.9
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Liquid
Vapor
Liquid
Vapor
0.2996 0.3021 0.3046 0.3074 0.3103 0.3134 0.3167 0.3203 0.3240 0.3281 0.3325 0.3372 0.3424 0.3481 0.3548 0.3627 0.3719 0.3829 0.3963 0.4131 0.4352 0.4659 0.5123 0.5929 0.7717 1.4786
0.1788 0.1829 0.1872 0.1918 0.1966 0.2017 0.2070 0.2127 0.2188 0.2253 0.2323 0.2398 0.2481 0.2572 0.2674 0.2790 0.2925 0.3083 0.3276 0.3520 0.3839 0.4286 0.4960 0.6112 0.8544 1.6683
0.0636 0.0626 0.0613 0.0602 0.0589 0.0576 0.0563 0.0550 0.0536 0.0522 0.0507 0.0492 0.0476 0.0460 0.0444 0.0427 0.0410 0.0392 0.0374 0.0355 0.0335 0.0314 0.0292 0.0267 0.0239 0.0199
0.00466 0.00497 0.00529 0.00559 0.00589 0.00619 0.00648 0.00676 0.00704 0.00732 0.00758 0.00785 0.00810 0.00835 0.00860 0.00884 0.00908 0.00931 0.00954 0.00976 0.00998 0.01020 0.01041 0.01063 0.01085 0.01110
Liquid 3.278 3.004 2.762 2.546 2.354 2.181 2.024 1.883 1.752 1.633 1.522 1.420 1.324 1.234 1.149 1.068 9.911 9.175 8.464 7.778 7.108 6.450 5.792 5.119 4.397 3.483
# # # # # # # # # # # # # # # # # # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
Vapor 1.714 2.053 2.433 2.856 3.314 3.811 4.342 4.906 5.494 6.103 6.725 7.356 7.986 8.611 9.222 9.814 1.038 1.092 1.144 1.195 1.245 1.298 1.356 1.431 1.544 1.787
Liquid
Vapor
Volume Expansion Coefficient b, 1/R Liquid
5.558 5.226 4.937 4.684 4.463 4.269 4.098 3.947 3.814 3.697 3.594 3.504 3.425 3.357 3.303 3.262 3.235 3.223 3.229 3.259 3.324 3.443 3.661 4.090 5.119 9.311
0.237 0.272 0.310 0.352 0.398 0.447 0.500 0.555 0.614 0.677 0.742 0.810 0.880 0.955 1.032 1.115 1.204 1.303 1.416 1.551 1.725 1.963 2.327 2.964 4.376 9.669
0.00114 0.00117 0.00120 0.00124 0.00128 0.00132 0.00137 0.00142 0.00149 0.00156 0.00163 0.00173 0.00183 0.00195 0.00210 0.00227 0.00248 0.00275 0.00308 0.00351 0.00411 0.00498 0.00637 0.00891 0.01490 0.04021
Prandtl Number Pr
Dynamic Viscosity m, lbm/ft · s
# # # # # # # # # # # # # # # # # # # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
Surface Tension, lbf/ft 0.001206 0.001146 0.001087 0.001029 0.000972 0.000915 0.000859 0.000803 0.000749 0.000695 0.000642 0.000590 0.000538 0.000488 0.000439 0.000391 0.000344 0.000299 0.000255 0.000212 0.000171 0.000132 0.000095 0.000061 0.000031 0.000006
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit Btu/lbm · °F for specific heat is equivalent to Btu/lbm · R, and the unit Btu/h · ft · °F for thermal conductivity is equivalent to Btu/h · ft · R. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: R. Tillner-Roth and H. D. Baehr, “An International Standard Formulation for the Thermodynamic Properties of 1,1,1,2-Tetrafluoroethane (HFC-134a) for Temperatures from 170 K to 455 K and Pressures up to 70 MPa,” J. Phys. Chem. Ref. Data, Vol. 23, No. 5, 1994; M. J. Assael, N. K. Dalaouti, A. A. Griva, and J. H. Dymond, “Viscosity and Thermal Conductivity of Halogenated Methane and Ethane Refrigerants,” IJR, Vol. 22, pp. 525–535, 1999; NIST REFPROP 6 program (M. O. McLinden, S. A. Klein, E. W. Lemmon, and A. P. Peskin, Physicial and Chemical Properties Division, National Institute of Standards and Technology, Boulder, CO 80303, 1995).
cen72367_appx2.qxd 11/17/04 4:34 PM Page 908
908 FLUID MECHANICS
TA B L E A – 5 E Properties of saturated ammonia Density r, lbm/ft3
Temp. T, °F
Saturation Pressure P, psia
Liquid
Vapor
"40 "30 "20 "10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240
10.4 13.9 18.3 23.7 30.4 38.5 48.2 59.8 73.4 89.2 107.7 128.9 153.2 180.8 212.0 247.2 286.5 330.4 379.2 433.2 492.7 558.2 630.1 708.6 794.4 887.9 989.5 1099.8 1219.4
43.08 42.66 42.33 41.79 41.34 40.89 40.43 39.96 39.48 38.99 38.50 37.99 37.47 36.94 36.40 35.83 35.26 34.66 34.04 33.39 32.72 32.01 31.26 30.47 29.62 28.70 27.69 25.57 25.28
0.0402 0.0527 0.0681 0.0869 0.1097 0.1370 0.1694 0.2075 0.2521 0.3040 0.3641 0.4332 0.5124 0.6029 0.7060 0.8233 0.9564 1.1074 1.2786 1.4730 1.6940 1.9460 2.2346 2.5670 2.9527 3.4053 3.9440 4.5987 5.4197
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Liquid
Vapor
Liquid
Vapor
Liquid
Vapor
Liquid
597.0 590.2 583.2 575.9 568.4 560.7 552.6 544.4 535.8 526.9 517.7 508.1 498.2 487.8 477.0 465.8 454.1 441.7 428.8 415.2 400.8 385.4 369.1 351.6 332.7 312.0 289.2 263.5 234.0
1.0542 1.0610 1.0677 1.0742 1.0807 1.0873 1.0941 1.1012 1.1087 1.1168 1.1256 1.1353 1.1461 1.1582 1.1719 1.1875 1.2054 1.2261 1.2502 1.2785 1.3120 1.3523 1.4015 1.4624 1.5397 1.6411 1.7798 1.9824 2.3100
0.5354 0.5457 0.5571 0.5698 0.5838 0.5992 0.6160 0.6344 0.6544 0.6762 0.6999 0.7257 0.7539 0.7846 0.8183 0.8554 0.8965 0.9425 0.9943 1.0533 1.1214 1.2012 1.2965 1.4128 1.5586 1.7473 2.0022 2.3659 2.9264
— — 0.3501 0.3426 0.3352 0.3278 0.3203 0.3129 0.3055 0.2980 0.2906 0.2832 0.2757 0.2683 0.2609 0.2535 0.2460 0.2386 0.2312 0.2237 0.2163 0.2089 0.2014 0.1940 0.1866 0.1791 0.1717 0.1643 0.1568
0.01026 0.01057 0.01089 0.01121 0.01154 0.01187 0.01220 0.01254 0.01288 0.01323 0.01358 0.01394 0.01431 0.01468 0.01505 0.01543 0.01582 0.01621 0.01661 0.01702 0.01744 0.01786 0.01829 0.01874 0.01919 0.01966 0.02015 0.02065 0.02119
1.966 # 10"4 1.853 # 10"4 1.746 # 10"4 1.645 # 10"4 1.549 # 10"4 1.458 # 10"4 1.371 # 10"4 1290 # 10"4 1.213 # 10"4 1.140 # 10"4 1.072 # 10"4 1.008 # 10"4 9.486 # 10"5 8.922 # 10"5 8.397 # 10"5 7.903 # 10"5 7.444 # 10"5 7.017 # 10"5 6.617 # 10"5 6.244 # 10"5 5.900 # 10"5 5.578 # 10"5 5.278 # 10"5 5.000 # 10"5 4.742 # 10"5 4.500 # 10"5 4.275 # 10"5 4.064 # 10"5 3.864 # 10"5
5.342 # 10"6 5.472 # 10"6 5.600 # 10"6 5.731 # 10"6 5.861 # 10"6 5.994 #10"6 6.125 #10"6 6.256 #10"6 6.389 #10"6 6.522 #10"6 6.656 #10"6 6.786 #10"6 6.922 #10"6 7.056 #10"6 7.189 #10"6 7.325 # 10"6 7.458 # 10"6 7.594 # 10"6 7.731 # 10"6 7.867 # 10"6 8.006 # 10"6 8.142 # 10"6 8.281 # 10"6 8.419 # 10"6 8.561 # 10"6 8.703 # 10"6 8.844 # 10"6 8.989 #10"6 9.136 #10"6
— — 1.917 1.856 1.797 1.740 1.686 1.634 1.585 1.539 1.495 1.456 1.419 1.387 1.358 1.333 1.313 1.298 1.288 1.285 1.288 1.300 1.322 1.357 1.409 1.484 1.595 1.765 2.049
Dynamic Viscosity m, lbm/ft · s
Vapor
Volume Expansion Coefficient b, 1/R Liquid
Surface Tension, lbf/ft
1.003 1.017 1.031 1.048 1.068 1.089 1.113 1.140 1.168 1.200 1.234 1.272 1.313 1.358 1.407 1.461 1.522 1.589 1.666 1.753 1.853 1.971 2.113 2.286 2.503 2.784 3.164 3.707 4.542
0.00098 0.00101 0.00103 0.00106 0.00109 0.00112 0.00116 0.00119 0.00123 0.00128 0.00132 0.00137 0.00143 0.00149 0.00156 0.00164 0.00174 0.00184 0.00196 0.00211 0.00228 0.00249 0.00274 0.00306 0.00348 0.00403 0.00480 0.00594 0.00784
0.002443 0.002357 0.002272 0.002187 0.002103 0.002018 0.001934 0.001850 0.001767 0.001684 0.001601 0.001518 0.001436 0.001354 0.001273 0.001192 0.001111 0.001031 0.000951 0.000872 0.000794 0.000716 0.000638 0.000562 0.000486 0.000411 0.000338 0.000265 0.000194
Prandtl Number Pr
Enthalpy of Vaporization hfg, Btu/lbm
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit Btu/lbm · °F for specific heat is equivalent to Btu/lbm · R, and the unit Btu/h · ft · °F for thermal conductivity is equivalent to Btu/h · ft · R. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Tillner-Roth, Harms-Watzenterg, and Baehr, “Eine neue Fundamentalgleichung fur Ammoniak,” DKV-Tagungsbericht 20: 167–181, 1993; Liley and Desai, “Thermophysical Properties of Refrigerants,” ASHRAE, 1993, ISBN 1-1883413-10-9.
cen72367_appx2.qxd 11/17/04 4:34 PM Page 909
909 APPENDIX 2
TA B L E A – 6 E Properties of saturated propane
Temp. T, °F "200 "180 "160 "140 "120 "100 "90 "80 "70 "60 "50 "40 "30 "20 "10 0 10 20 30 40 50 60 70 80 90 100 120 140 160 180
Saturation Pressure P, psia 0.0201 0.0752 0.2307 0.6037 1.389 2.878 4.006 5.467 7.327 9.657 12.54 16.05 20.29 25.34 31.3 38.28 46.38 55.7 66.35 78.45 92.12 107.5 124.6 143.7 164.8 188.1 241.8 306.1 382.4 472.9
Density r, lbm/ft3 Liquid
Vapor
42.06 41.36 40.65 39.93 39.20 38.46 38.08 37.70 37.32 36.93 36.54 36.13 35.73 35.31 34.89 34.46 34.02 33.56 33.10 32.62 32.13 31.63 31.11 30.56 30.00 29.41 28.13 26.69 24.98 22.79
0.0003 0.0011 0.0032 0.0078 0.0170 0.0334 0.0453 0.0605 0.0793 0.1024 0.1305 0.1641 0.2041 0.2512 0.3063 0.3703 0.4441 0.5289 0.6259 0.7365 0.8621 1.0046 1.1659 1.3484 1.5549 1.7887 2.3562 3.1003 4.1145 5.6265
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Enthalpy of Vaporization hfg, Btu/lbm
Liquid
Vapor
Liquid
Vapor
217.7 213.4 209.1 204.8 200.5 196.1 193.9 191.6 189.3 186.9 184.4 181.9 179.3 176.6 173.8 170.9 167.9 164.8 161.6 158.1 154.6 150.8 146.8 142.7 138.2 133.6 123.2 111.1 96.4 77.1
0.4750 0.4793 0.4845 0.4907 0.4982 0.5069 0.5117 0.5169 0.5224 0.5283 0.5345 0.5392 0.5460 0.5531 0.5607 0.5689 0.5775 0.5867 0.5966 0.6072 0.6187 0.6311 0.6447 0.6596 0.6762 0.6947 0.7403 0.7841 0.8696 1.1436
0.2595 0.2680 0.2769 0.2866 0.2971 0.3087 0.3150 0.3215 0.3284 0.3357 0.3433 0.3513 0.3596 0.3684 0.3776 0.3874 0.3976 0.4084 0.4199 0.4321 0.4452 0.4593 0.4746 0.4915 0.5103 0.5315 0.5844 0.6613 0.7911 1.0813
0.1073 0.1033 0.0992 0.0949 0.0906 0.0863 0.0842 0.0821 0.0800 0.0780 0.0760 0.0740 0.0721 0.0702 0.0683 0.0665 0.0647 0.0629 0.0612 0.0595 0.0579 0.0563 0.0547 0.0532 0.0517 0.0501 0.0472 0.0442 0.0411 0.0376
0.00313 0.00347 0.00384 0.00423 0.00465 0.00511 0.00534 0.00559 0.00585 0.00611 0.00639 0.00668 0.00697 0.00728 0.00761 0.00794 0.00829 0.00865 0.00903 0.00942 0.00983 0.01025 0.01070 0.01116 0.01165 0.01217 0.01328 0.01454 0.01603 0.01793
Liquid 5.012 # 3.941 # 3.199 # 2.660 # 2.252 # 1.934 # 1.799 # 1.678 # 1.569 # 1.469 # 1.378 # 1.294 # 1.217 # 1.146 # 1.079 # 1.018 # 9.606 # 9.067 # 8.561 # 8.081 # 7.631 # 7.200 # 6.794 # 6.406 # 6.033 # 5.675 # 5.000 # 4.358 # 3.733 # 3.083 #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
Vapor 2.789 2.975 3.164 3.358 3.556 3.756 3.858 3.961 4.067 4.172 4.278 4.386 4.497 4.611 4.725 4.842 4.961 5.086 5.211 5.342 5.478 5.617 5.764 5.919 6.081 6.256 6.644 7.111 7.719 8.617
Liquid
Vapor
Volume Expansion Coefficient b, 1/R Liquid
7.991 6.582 5.626 4.951 4.457 4.087 3.936 3.803 3.686 3.582 3.490 3.395 3.320 3.253 3.192 3.137 3.088 3.043 3.003 2.967 2.935 2.906 2.881 2.860 2.843 2.831 2.825 2.784 2.845 3.380
0.833 0.826 0.821 0.818 0.817 0.817 0.819 0.820 0.822 0.825 0.828 0.831 0.835 0.840 0.845 0.850 0.857 0.864 0.873 0.882 0.893 0.906 0.921 0.938 0.959 0.984 1.052 1.164 1.371 1.870
0.00083 0.00086 0.00088 0.00091 0.00094 0.00097 0.00099 0.00101 0.00104 0.00106 0.00109 0.00112 0.00115 0.00119 0.00123 0.00127 0.00132 0.00138 0.00144 0.00151 0.00159 0.00168 0.00179 0.00191 0.00205 0.00222 0.00267 0.00338 0.00459 0.00791
Prandtl Number Pr
Dynamic Viscosity m, lbm/ft · s
# # # # # # # # # # # # # # # # # # # # # # # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6
Surface Tension, lbf/ft 0.001890 0.001780 0.001671 0.001563 0.001455 0.001349 0.001297 0.001244 0.001192 0.001140 0.001089 0.001038 0.000987 0.000937 0.000887 0.000838 0.000789 0.000740 0.000692 0.000644 0.000597 0.000551 0.000505 0.000460 0.000416 0.000372 0.000288 0.000208 0.000133 0.000065
Note 1: Kinematic viscosity n and thermal diffusivity a can be calculated from their definitions, n ! m/r and a ! k/rcp ! n/Pr. The properties listed here (except the vapor density) can be used at any pressures with negligible error except at temperatures near the critical-point value. Note 2: The unit Btu/lbm · °F for specific heat is equivalent to Btu/lbm · R, and the unit Btu/h · ft · °F for thermal conductivity is equivalent to Btu/h · ft · R. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Reiner Tillner-Roth, “Fundamental Equations of State,” Shaker, Verlag, Aachan, 1998; B. A. Younglove and J. F. Ely, “Thermophysical Properties of Fluids. II Methane, Ethane, Propane, Isobutane, and Normal Butane,” J. Phys. Chem. Ref. Data, Vol. 16, No. 4, 1987; G. R. Somayajulu, “A Generalized Equation for Surface Tension from the Triple-Point to the CriticalPoint,” International Journal of Thermophysics, Vol. 9, No. 4, 1988.
cen72367_appx2.qxd 11/17/04 4:34 PM Page 910
910 FLUID MECHANICS
TA B L E A – 7 E Properties of liquids
Temp. T, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Thermal Diffusivity a, ft2/s
Dynamic Viscosity m, lbm/ft · s
Kinematic Viscosity n, ft2/s
Prandtl Number Pr
Volume Expansion Coeff. b, 1/R
Methane (CH4) "280 "260 "240 "220 "200 "180 "160 "140
27.41 26.43 25.39 24.27 23.04 21.64 19.99 17.84
0.8152 0.8301 0.8523 0.8838 0.9314 1.010 1.158 1.542
0.1205 0.1097 0.0994 0.0896 0.0801 0.0709 0.0616 0.0518
1.497 1.389 1.276 1.159 1.036 9.008 7.397 5.234
# # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"7 10"7 10"7
1.057 8.014 6.303 5.075 4.142 3.394 2.758 2.168
# # # # # # # #
10"4 10"5 10"5 10"5 10"5 10"5 10"5 10"5
3.857 3.032 2.482 2.091 1.798 1.568 1.379 1.215
# # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"6
2.575 2.183 1.945 1.803 1.734 1.741 1.865 2.322
0.00175 0.00192 0.00215 0.00247 0.00295 0.00374 0.00526 0.00943
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
7.879 6.840 6.005 5.326 4.769 4.308
# # # # # #
10"6 10"6 10"6 10"6 10"6 10"6
7.317 6.468 5.793 5.250 4.808 4.447
0.000656 0.000671 0.000691 0.000716 0.000749 0.000789
# # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"5 10"5 10"5
1.500 8.939 6.043 4.406 3.368 2.653 2.127 1.713
# # # # # # # #
10"5 10"6 10"6 10"6 10"6 10"6 10"6 10"6
12.96 7.977 5.830 4.738 4.114 3.716 3.437 3.264
0.000785 0.000836 0.000908 0.001012 0.001169 0.001421 0.001883 0.002970
Methanol [CH3(OH)] 70 90 110 130 150 170
49.15 48.50 47.85 47.18 46.50 45.80
0.6024 0.6189 0.6373 0.6576 0.6796 0.7035
0.1148 0.1143 0.1138 0.1133 0.1128 0.1124
1.076 1.057 1.036 1.014 9.918 9.687
# # # # # #
10"6 10"6 10"6 10"6 10"7 10"7
3.872 3.317 2.872 2.513 2.218 1.973
Isobutane (R600a) "150 "100 "50 0 50 100 150 200
42.75 41.06 39.31 37.48 35.52 33.35 30.84 27.73
0.4483 0.4721 0.4986 0.5289 0.5643 0.6075 0.6656 0.7635
0.0799 0.0782 0.0731 0.0664 0.0591 0.0521 0.0457 0.0400
1.157 # 1.120 # 1.036 # 9.299 # 8.187 # 7.139 # 6.188 # 5.249 #
10"6 10"6 10"6 10"7 10"7 10"7 10"7 10"7
6.417 3.669 2.376 1.651 1.196 8.847 6.558 4.750
Glycerin 32 40 50 60 70 80 90 100
79.65 79.49 79.28 79.07 78.86 78.66 78.45 78.24
0.5402 0.5458 0.5541 0.5632 0.5715 0.5794 0.5878 0.5964
0.163 0.1637 0.1645 0.1651 0.1652 0.1652 0.1652 0.1653
1.052 1.048 1.040 1.029 1.018 1.007 9.955 9.841
# # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"7 10"7
7.047 4.803 2.850 1.547 0.9422 0.5497 0.3756 0.2277
0.08847 0.06042 0.03594 0.01956 0.01195 0.00699 0.004787 0.00291
84101 57655 34561 18995 11730 6941 4809 2957
Engine Oil (unused) 32 50 75 100 125 150 200 250 300
56.12 55.79 55.3 54.77 54.24 53.73 52.68 51.71 50.63
0.4291 0.4395 0.4531 0.4669 0.4809 0.4946 0.5231 0.5523 0.5818
0.0849 0.08338 0.08378 0.08367 0.08207 0.08046 0.07936 0.07776 0.07673
9.792 9.448 9.288 9.089 8.740 8.411 7.999 7.563 7.236
# # # # # # # # #
10"7 10"7 10"7 10"7 10"7 10"7 10"7 10"7 10"7
2.563 1.210 0.4286 0.1630 7.617 # 3.833 # 1.405 # 6.744 # 3.661 #
10"2 10"2 10"2 10"3 10"3
4.566 2.169 7.751 2.977 1.404 7.135 2.668 1.304 7.232
# # # # # # # # #
10"2 46636 10"2 22963 10"3 8345 10"3 3275 10"3 1607 10"4 848.3 10"4 333.6 10"4 172.5 10"5 99.94
Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
0.000389 0.000389 0.000389 0.000389 0.000389 0.000389 0.000389 0.000389 0.000389
cen72367_appx2.qxd 11/17/04 4:34 PM Page 911
911 APPENDIX 2
TA B L E A – 8 E Properties of liquid metals
Temp. T, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Thermal Diffusivity a, ft2/s
Dynamic Viscosity m, lbm/ft · s
Kinematic Viscosity n, ft2/s
Prandtl Number Pr
Volume Expansion Coeff. b, 1/R
Mercury (Hg) Melting Point: "38°F 32 50 100 150 200 300 400 500 600
848.7 847.2 842.9 838.7 834.5 826.2 817.9 809.6 801.3
0.03353 0.03344 0.03319 0.03298 0.03279 0.03252 0.03236 0.03230 0.03235
4.727 4.805 5.015 5.221 5.422 5.815 6.184 6.518 6.839
4.614 4.712 4.980 5.244 5.504 6.013 6.491 6.924 7.329
# # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
1.133 1.092 9.919 9.122 8.492 7.583 6.972 6.525 6.186
# # # # # # # # #
10"3 10"3 10"4 10"4 10"4 10"4 10"4 10"4 10"4
1.335 1.289 1.176 1.087 1.017 9.180 8.524 8.061 7.719
# # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"7 10"7 10"7 10"7
0.02895 0.02737 0.02363 0.02074 0.01849 0.01527 0.01313 0.01164 0.01053
1.614 1.482 1.350 1.215 1.138
# # # # #
10"6 10"6 10"6 10"6 10"6
0.01352 0.01271 0.01183 0.0109 0.01029
2.450 2.223 1.994 1.862 1.727 1.590
# # # # # #
10"6 10"6 10"6 10"6 10"6 10"6
0.02369 0.02143 0.01917 0.01798 0.01676 0.01551
7.239 6.350 5.433 4.488 3.354 3.014
# # # # # #
10"6 10"6 10"6 10"6 10"6 10"6
0.01007 0.008891 0.007667 0.006387 0.004860 0.004449
4.933 4.500 4.052 3.589 2.614 2.409
# # # # # #
10"6 10"6 10"6 10"6 10"6 10"6
0.006577 0.005975 0.005359 0.004728 0.003420 0.003248
# # # # # #
10"6 10"6 10"6 10"6 10"6 10"6
0.02136 0.01857 0.0159 0.01095 0.007296 0.006324
Bismuth (Bi) Melting Point: 520°F 700 800 900 1000 1100
620.7 616.5 612.2 608.0 603.7
0.03509 0.03569 0.0363 0.0369 0.0375
9.361 9.245 9.129 9.014 9.014
1.193 1.167 1.141 1.116 1.105
# # # # #
10"4 10"4 10"4 10"4 10"4
1.001 9.142 8.267 7.392 6.872
# # # # #
10"3 10"4 10"4 10"4 10"4
Lead (Pb) Melting Point: 621°F 700 800 900 1000 1100 1200
658 654 650 645.7 641.5 637.2
0.03797 0.03750 0.03702 0.03702 0.03702 0.03702
9.302 9.157 9.013 8.912 8.810 8.709
1.034 1.037 1.040 1.035 1.030 1.025
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
1.612 1.453 1.296 1.202 1.108 1.013
# # # # # #
10"3 10"3 10"3 10"3 10"3 10"3
Sodium (Na) Melting Point: 208°F 300 400 500 600 800 1000
57.13 56.28 55.42 54.56 52.85 51.14
0.3258 0.3219 0.3181 0.3143 0.3089 0.3057
48.19 46.58 44.98 43.37 40.55 38.12
7.192 7.142 7.087 7.026 6.901 6.773
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
4.136 3.572 3.011 2.448 1.772 1.541
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
Potassium (K) Melting Point: 147°F 300 400 500 600 800 1000
50.40 49.58 48.76 47.94 46.31 44.62
0.1911 0.1887 0.1863 0.1839 0.1791 0.1791
26.00 25.37 24.73 24.09 22.82 21.34
7.500 7.532 7.562 7.591 7.643 7.417
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
2.486 2.231 1.976 1.721 1.210 1.075
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
Sodium–Potassium (%22Na–%78K) Melting Point: 12°F 200 300 400 600 800 1000
52.99 52.16 51.32 49.65 47.99 46.36
0.2259 0.2230 0.2201 0.2143 0.2100 0.2103
14.79 14.99 15.19 15.59 15.95 16.20
3.432 3.580 3.735 4.070 4.396 4.615
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
3.886 3.467 3.050 2.213 1.539 1.353
# # # # # #
10"4 10"4 10"4 10"4 10"4 10"4
7.331 6.647 5.940 4.456 3.207 2.919
Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
1.005 1.005 1.005 1.005 1.005 1.005 1.008 1.018 1.035
# # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4
cen72367_appx2.qxd 11/17/04 4:34 PM Page 912
912 FLUID MECHANICS
TA B L E A – 9 E Properties of air at 1 atm pressure Temp. T, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
"300 "200 "100 "50 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 250 300 350 400 450 500 600 700 800 900 1000 1500 2000 2500 3000 3500 4000
0.24844 0.15276 0.11029 0.09683 0.08630 0.08446 0.08270 0.08101 0.07939 0.07783 0.07633 0.07489 0.07350 0.07217 0.07088 0.06963 0.06843 0.06727 0.06615 0.06507 0.06402 0.06300 0.06201 0.06106 0.06013 0.05590 0.05222 0.04899 0.04614 0.04361 0.04134 0.03743 0.03421 0.03149 0.02917 0.02718 0.02024 0.01613 0.01340 0.01147 0.01002 0.00889
0.5072 0.2247 0.2360 0.2389 0.2401 0.2402 0.2403 0.2403 0.2404 0.2404 0.2404 0.2404 0.2404 0.2404 0.2405 0.2405 0.2405 0.2405 0.2406 0.2406 0.2406 0.2407 0.2408 0.2408 0.2409 0.2415 0.2423 0.2433 0.2445 0.2458 0.2472 0.2503 0.2535 0.2568 0.2599 0.2630 0.2761 0.2855 0.2922 0.2972 0.3010 0.3040
0.00508 0.00778 0.01037 0.01164 0.01288 0.01312 0.01336 0.01361 0.01385 0.01409 0.01433 0.01457 0.01481 0.01505 0.01529 0.01552 0.01576 0.01599 0.01623 0.01646 0.01669 0.01692 0.01715 0.01738 0.01761 0.01874 0.01985 0.02094 0.02200 0.02305 0.02408 0.02608 0.02800 0.02986 0.03164 0.03336 0.04106 0.04752 0.05309 0.05811 0.06293 0.06789
Thermal Diffusivity a, ft2/s 1.119 6.294 1.106 1.397 1.726 1.797 1.868 1.942 2.016 2.092 2.169 2.248 2.328 2.409 2.491 2.575 2.660 2.746 2.833 2.921 3.010 3.100 3.191 3.284 3.377 3.857 4.358 4.879 5.419 5.974 6.546 7.732 8.970 1.025 1.158 1.296 2.041 2.867 3.765 4.737 5.797 6.975
# # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # #
10"5 10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3
Dynamic Viscosity m, lbm/ft · s 4.039 6.772 9.042 1.006 1.102 1.121 1.140 1.158 1.176 1.194 1.212 1.230 1.247 1.265 1.281 1.299 1.316 1.332 1.349 1.365 1.382 1.398 1.414 1.430 1.446 1.524 1.599 1.672 1.743 1.812 1.878 2.007 2.129 2.247 2.359 2.467 2.957 3.379 3.750 4.082 4.381 4.651
# # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # #
10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
Kinematic Viscosity n, ft2/s 1.625 4.433 8.197 1.039 1.278 1.328 1.379 1.430 1.482 1.535 1.588 1.643 1.697 1.753 1.809 1.866 1.923 1.981 2.040 2.099 2.159 2.220 2.281 2.343 2.406 2.727 3.063 3.413 3.777 4.154 4.544 5.361 6.225 7.134 8.087 9.080 1.460 2.095 2.798 3.560 4.373 5.229
# # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # # #
10"5 10"5 10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3 10"3 10"3 10"3
Prandtl Number Pr 1.4501 0.7042 0.7404 0.7439 0.7403 0.7391 0.7378 0.7365 0.7350 0.7336 0.7321 0.7306 0.7290 0.7275 0.7260 0.7245 0.7230 0.7216 0.7202 0.7188 0.7174 0.7161 0.7148 0.7136 0.7124 0.7071 0.7028 0.6995 0.6971 0.6953 0.6942 0.6934 0.6940 0.6956 0.6978 0.7004 0.7158 0.7308 0.7432 0.7516 0.7543 0.7497
Note: For ideal gases, the properties cp, k, m, and Pr are independent of pressure. The properties r, n, and a at a pressure P (in atm) other than 1 atm are determined by multiplying the values of r at the given temperature by P and by dividing n and a by P. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Original sources: Keenan, Chao, Keyes, Gas Tables, Wiley, 198; and Thermophysical Properties of Matter, Vol. 3: Thermal Conductivity, Y. S. Touloukian, P. E. Liley, S. C. Saxena, Vol. 11: Viscosity, Y. S. Touloukian, S. C. Saxena, and P. Hestermans, IFI/Plenun, NY, 1970, ISBN 0-306067020-8.
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913 APPENDIX 2
TA B L E A – 1 0 E Properties of gases at 1 atm pressure Temp. T, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Thermal Diffusivity a, ft2/s
Dynamic Viscosity m, lbm/ft · s
Kinematic Viscosity n, ft2/s
Prandtl Number Pr
Carbon Dioxide, CO2 "50 0 50 100 200 300 500 1000 1500 2000
0.14712 0.13111 0.11825 0.10769 0.09136 0.07934 0.06280 0.04129 0.03075 0.02450
0.1797 0.1885 0.1965 0.2039 0.2171 0.2284 0.2473 0.2796 0.2995 0.3124
0.00628 0.00758 0.00888 0.01017 0.01273 0.01528 0.02027 0.03213 0.04281 0.05193
6.600 8.522 1.061 1.286 1.784 2.341 3.626 7.733 1.290 1.885
# # # # # # # # # #
10"5 10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3
7.739 8.661 9.564 1.045 1.217 1.382 1.696 2.381 2.956 3.451
# # # # # # # # # #
10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5
5.261 6.606 8.086 9.703 1.332 1.743 2.700 5.767 9.610 1.408
# # # # # # # # # #
10"5 10"5 10"5 10"5 10"4 10"4 10"4 10"4 10"4 10"3
0.7970 0.7751 0.7621 0.7543 0.7469 0.7445 0.7446 0.7458 0.7445 0.7474
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
9.419 1.036 1.127 1.214 1.379 1.531 1.802 2.334 2.766 3.231
# # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
1.005 1.242 1.498 1.772 2.372 3.032 4.508 8.881 1.413 2.072
# # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3
0.7798 0.7593 0.7454 0.7359 0.7247 0.7191 0.7143 0.7078 0.7038 0.7136
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
5.861 6.506 7.133 7.742 8.906 1.000 1.200 1.620 1.974 2.327
# # # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"5 10"5 10"5 10"5 10"5
1.092 1.361 1.655 1.972 2.674 3.457 5.244 1.076 1.760 2.605
# # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
0.8033 0.7649 0.7428 0.7311 0.7245 0.7283 0.7412 0.7491 0.7366 0.7353
10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"3 10"2 10"2
4.969 5.381 5.781 6.167 6.911 7.622 8.967 1.201 1.477 1.734
# # # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"6 10"5 10"5 10"5
7.373 8.960 1.067 1.250 1.652 2.098 3.117 6.354 1.048 1.544
# # # # # # # # # #
10"4 10"4 10"3 10"3 10"3 10"3 10"3 10"3 10"2 10"2
0.6638 0.6960 0.7112 0.7177 0.7197 0.7174 0.7146 0.7241 0.7323 0.7362
Carbon Monoxide, CO "50 0 50 100 200 300 500 1000 1500 2000
0.09363 0.08345 0.07526 0.06854 0.05815 0.05049 0.03997 0.02628 0.01957 0.01559
0.2571 0.2523 0.2496 0.2484 0.2485 0.2505 0.2567 0.2732 0.2862 0.2958
0.01118 0.01240 0.01359 0.01476 0.01702 0.01920 0.02331 0.03243 0.04049 0.04822
1.290 1.636 2.009 2.408 3.273 4.217 6.311 1.254 2.008 2.903
# # # # # # # # # #
Methane, CH4 "50 0 50 100 200 300 500 1000 1500 2000
0.05363 0.04779 0.04311 0.03925 0.03330 0.02892 0.02289 0.01505 0.01121 0.00893
0.5335 0.5277 0.5320 0.5433 0.5784 0.6226 0.7194 0.9438 1.1162 1.2419
0.01401 0.01616 0.01839 0.02071 0.02559 0.03077 0.04195 0.07346 0.10766 0.14151
1.360 1.780 2.228 2.698 3.690 4.748 7.075 1.436 2.390 3.544
# # # # # # # # # #
Hydrogen, H2 "50 0 50 100 200 300 500 1000 1500 2000
0.00674 0.00601 0.00542 0.00493 0.00419 0.00363 0.00288 0.00189 0.00141 0.00112
3.0603 3.2508 3.3553 3.4118 3.4549 3.4613 3.4572 3.5127 3.6317 3.7656
0.08246 0.09049 0.09818 0.10555 0.11946 0.13241 0.15620 0.20989 0.26381 0.31923
1.110 1.287 1.500 1.742 2.295 2.924 4.363 8.776 1.432 2.098
# # # # # # # # # #
(Continued)
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914 FLUID MECHANICS
TA B L E A – 1 0 E Properties of gases at 1 atm pressure (Continued) Temp. T, °F
Density r, lbm/ft3
Specific Heat cp, Btu/lbm · R
Thermal Conductivity k, Btu/h · ft · R
Thermal Diffusivity a, ft2/s
Dynamic Viscosity m, lbm/ft · s
Kinematic Viscosity n, ft2/s
Prandtl Number Pr
Nitrogen, N2 "50 0 50 100 200 300 500 1000 1500 2000
0.09364 0.08346 0.07527 0.06854 0.05815 0.05050 0.03997 0.02628 0.01958 0.01560
0.2320 0.2441 0.2480 0.2489 0.2487 0.2492 0.2535 0.2697 0.2831 0.2927
0.01176 0.01300 0.01420 0.01537 0.01760 0.01970 0.02359 0.03204 0.04002 0.04918
1.504 1.773 2.113 2.502 3.379 4.349 6.466 1.255 2.006 2.992
# # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
9.500 1.043 1.134 1.221 1.388 1.543 1.823 2.387 2.829 3.212
# # # # # # # # # #
10"6 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
1.014 1.251 1.507 1.783 2.387 3.055 4.559 9.083 1.445 2.059
# # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3
0.6746 0.7056 0.7133 0.7126 0.7062 0.7025 0.7051 0.7232 0.7202 0.6882
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
1.104 1.218 1.326 1.429 1.625 1.806 2.139 2.855 3.474 4.035
# # # # # # # # # #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
1.032 1.277 1.543 1.826 2.446 3.132 4.685 9.509 1.553 2.265
# # # # # # # # # #
10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3
0.7622 0.7312 0.7152 0.7072 0.7020 0.7018 0.7029 0.7005 0.6985 0.6988
4.933 5.592 6.261 6.942 8.333 9.756 1.267 2.014 2.742 3.422
# # # # # # # # # #
10"6 10"6 10"6 10"6 10"6 10"6 10"5 10"5 10"5 10"5
8.192 1.041 1.293 1.574 2.228 3.004 4.931 1.191 2.178 3.411
# # # # # # # # # #
10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
1.0050 1.0049 1.0018 0.9969 0.9845 0.9713 0.9475 0.9063 0.8793 0.8563
Oxygen, O2 "50 0 50 100 200 300 500 1000 1500 2000
0.10697 0.09533 0.08598 0.07830 0.06643 0.05768 0.04566 0.03002 0.02236 0.01782
0.2331 0.2245 0.2209 0.2200 0.2221 0.2262 0.2352 0.2520 0.2626 0.2701
0.01216 0.01346 0.01475 0.01601 0.01851 0.02096 0.02577 0.03698 0.04701 0.05614
1.355 1.747 2.157 2.582 3.484 4.463 6.665 1.357 2.224 3.241
# # # # # # # # # #
Water Vapor, H2O "50 0 50 100 200 300 500 1000 1500 2000
0.06022 0.05367 0.04841 0.04408 0.03740 0.03248 0.02571 0.01690 0.01259 0.01003
0.4512 0.4484 0.4472 0.4473 0.4503 0.4557 0.4707 0.5167 0.5625 0.6034
0.00797 0.00898 0.01006 0.01121 0.01372 0.01648 0.02267 0.04134 0.06315 0.08681
8.153 1.036 1.291 1.579 2.263 3.093 5.204 1.314 2.477 3.984
# # # # # # # # # #
10"5 10"4 10"4 10"4 10"4 10"4 10"4 10"3 10"3 10"3
Note: For ideal gases, the properties cp, k, m, and Pr are independent of pressure. The properties r, n, and a at a pressure P (in atm) other than 1 atm are determined by multiplying the values of r at the given temperature by P and by dividing n and a by P. Source: Data generated from the EES software developed by S. A. Klein and F. L. Alvarado. Originally based on various sources.
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915 APPENDIX 2
TA B L E A – 1 1 E Properties of the atmosphere at high altitude Altitude, ft
Temperature, °F
Pressure, psia
Gravity, g, ft/s2
Speed of Sound, ft/s
Density, lbm/ft3
0 500 1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500 7000 7500 8000 8500 9000 9500 10,000 11,000 12,000 13,000 14,000 15,000 16,000 17,000 18,000 19,000 20,000 22,000 24,000 26,000 28,000 30,000 32,000 34,000 36,000 38,000 40,000 45,000 50,000 55,000 60,000
59.00 57.22 55.43 53.65 51.87 50.09 48.30 46.52 44.74 42.96 41.17 39.39 37.61 35.83 34.05 32.26 30.48 28.70 26.92 25.14 23.36 19.79 16.23 12.67 9.12 5.55 %1.99 "1.58 "5.14 "8.70 "12.2 "19.4 "26.5 "33.6 "40.7 "47.8 "54.9 "62.0 "69.2 "69.7 "69.7 "69.7 "69.7 "69.7 "69.7
14.7 14.4 14.2 13.9 13.7 13.4 13.2 12.9 12.7 12.5 12.2 12.0 11.8 11.6 11.3 11.1 10.9 10.7 10.5 10.3 10.1 9.72 9.34 8.99 8.63 8.29 7.97 7.65 7.34 7.05 6.76 6.21 5.70 5.22 4.78 4.37 3.99 3.63 3.30 3.05 2.73 2.148 1.691 1.332 1.048
32.174 32.173 32.171 32.169 32.168 32.166 32.165 32.163 32.162 32.160 32.159 32.157 32.156 32.154 32.152 32.151 32.149 32.148 32.146 32.145 32.145 32.140 32.137 32.134 32.131 32.128 32.125 32.122 32.119 32.115 32.112 32.106 32.100 32.094 32.088 32.082 32.08 32.07 32.06 32.06 32.05 32.04 32.02 32.00 31.99
1116 1115 1113 1111 1109 1107 1105 1103 1101 1099 1097 1095 1093 1091 1089 1087 1085 1083 1081 1079 1077 1073 1069 1065 1061 1057 1053 1049 1045 1041 1037 1029 1020 1012 1003 995 987 978 969 968 968 968 968 968 968
0.07647 0.07536 0.07426 0.07317 0.07210 0.07104 0.06998 0.06985 0.06792 0.06690 0.06590 0.06491 0.06393 0.06296 0.06200 0.06105 0.06012 0.05919 0.05828 0.05738 0.05648 0.05473 0.05302 0.05135 0.04973 0.04814 0.04659 0.04508 0.04361 0.04217 0.04077 0.03808 0.03553 0.03311 0.03082 0.02866 0.02661 0.02468 0.02285 0.02079 0.01890 0.01487 0.01171 0.00922 0.00726
Viscosity m, lbm/ft · s 1.202 # 1.199 # 1.196 # 1.193 # 1.190 # 1.186 # 1.183 # 1.180 # 1.177 # 1.173 # 1.170 # 1.167 # 1.164 # 1.160 # 1.157 # 1.154 # 1.150 # 1.147 # 1.144 # 1.140 # 1.137 # 1.130 # 1.124 # 1.117 # 1.110 # 1.104 # 1.097 # 1.090 # 1.083 # 1.076 # 1.070 # 1.056 # 1.042 # 1.028 # 1.014 # 1.000 # 0.986 # 0.971 # 0.956 # 0.955 # 0.955 # 0.955 # 0.955 # 0.955 # 0.955 #
10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5 10"5
Thermal Conductivity, Btu/h · ft · R 0.0146 0.0146 0.0146 0.0145 0.0145 0.0144 0.0144 0.0143 0.0143 0.0142 0.0142 0.0141 0.0141 0.0141 0.0140 0.0140 0.0139 0.0139 0.0138 0.0138 0.0137 0.0136 0.0136 0.0135 0.0134 0.0133 0.0132 0.0132 0.0130 0.0129 0.0128 0.0126 0.0124 0.0122 0.0121 0.0119 0.0117 0.0115 0.0113 0.0113 0.0113 0.0113 0.0113 0.0113 0.0113
Source: U.S. Standard Atmosphere Supplements, U.S. Government Printing Office, 1966. Based on year-round mean conditions at 45° latitude and varies with the time of the year and the weather patterns. The conditions at sea level (z ! D) are taken to be P ! 14.696 psia, T ! 59°F, r ! 0.076474 lbm/ft3, g ! 32.1741 ft2/s.
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cen72367_glos.qxd 11/23/04 11:09 AM Page 917
GLOSSARY Guest Author: James G. Brasseur, The Pennsylvania State University Note: Boldface color glossary terms correspond to boldface color terms in the text. The number in [square brackets] denotes the page of the boldface color term in the text. Italics indicates a term defined elsewhere in the glossary. Boldface terms without page numbers are concepts that are not defined in the text but are defined or cross-referenced in the glossary for students to review. absolute pressure [66]: See stress, pressure stress. Contrast with gage pressure. absolute viscosity [47]: See viscosity. acceleration field [122]: See field. adiabatic process [202]: A process with no heat transfer. advective acceleration [126]: In order to reduce confusion of terminology in flows where buoyancy forces generate convective fluid motions, the term “convective acceleration” is often replaced with the term “advective acceleration.” aerodynamics [2]: The application of fluid dynamics to air, land, and water-going vehicles. Often the term is specifically applied to the flow surrounding, and forces and moments on, flight vehicles in air, as opposed to vehicles in water or other liquids (hydrodynamics). angle of attack [570]: The angle between an airfoil or wing and the free-stream flow velocity vector. average: An area/volume/time average of a fluid property is the integral of the property over an area/volume/time period divided by the corresponding area/volume/time period. Also called mean. axisymmetric flow [419, 490, 492, 565]: A flow that when specified appropriately using cylindrical coordinates (r, u, x) does not vary in the azimuthal (u) direction. Thus, all partial derivatives in u are zero. The flow is therefore either onedimensional or two-dimensional (see also dimensionality and planar flow). barometer [75]: A device that measures atmospheric pressure. basic dimensions: See dimensions. Bernoulli equation [185, 187, 208, 270]: A useful reduction of conservation of momentum (and conservation of energy) that describes a balance between pressure (flow work), velocity (kinetic energy), and position of fluid particles relative to the gravity vector (potential energy) in regions
Note: This glossary covers boldface color terms found in Chapters 1 to 11.
of a fluid flow where frictional force on fluid particles is negligible compared to pressure force in that region of the flow (see inviscid flow). There are multiple forms of the Bernoulli equation for incompressible vs. compressible, steady vs. nonsteady, and derivations through Newton’s law vs. the first law of thermodynamics. The most commonly used forms are for steady incompressible fluid flow derived through conservation of momentum. bluff (or blunt) body [563]: A moving object with a blunt rear portion. Bluff bodies have wakes resulting from massive flow separation over the rear of the body. boundary condition [400, 440]: In solving for flow field variables (velocity, temperature) from governing equations, it is necessary to mathematically specify a function of the variable at the surface. These mathematical statements are called boundary conditions. The no-slip condition that the flow velocity must equal the surface velocity at the surface is an example of a boundary condition that is used with the Navier–Stokes equation to solve for the velocity field. boundary layer [6, 325, 481, 511]: At high Reynolds numbers relatively thin “boundary layers” exist in the flow adjacent to surfaces where the flow is brought to rest (see noslip condition). Boundary layers are characterized by high shear with the highest velocities away from the surface. Frictional force, viscous stress, and vorticity are significant in boundary layers. The approximate form of the two components of the Navier– Stokes equation, simplified by neglecting the terms that are small within the boundary layer, are called the boundary layer equations. The associated approximation based on the existence of thin boundary layers surrounded by irrotational or inviscid flow is called the boundary layer approximation. boundary layer approximation [511]: See boundary layer. boundary layer equations [511, 515]: See boundary layer. boundary layer thickness measures: Different measures of the thickness of a boundary layer as a function of downstream distance are used in fluid flow analyses. These are: boundary layer thickness [512]: The full thickness of the viscous layer that defines the boundary layer, from the surface to the edge. Defining the edge is difficult to do precisely, so the “edge” of the boundary layer is often defined as the point where the boundary layer velocity is a large fraction of the free-stream velocity (e.g., d99 is the distance from the surface to the point where the streamwise velocity component is 99 percent of the freestream velocity). displacement thickness [524]: A boundary layer thickness measure that quantifies the deflection of fluid
917
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918 FLUID MECHANICS
streamlines in the direction away from the surface as a result of friction-induced reduction in mass flow adjacent to the surface. Displacement thickness (d*) is a measure of the thickness of this mass flow rate deficit layer. In all boundary layers, d* d. momentum thickness [527]: A measure of the layer of highest deficit in momentum flow rate adjacent to the surface as a result of frictional resisting force (shear stress). Because Newton’s second law states that force equals time rate of momentum change, momentum thickness u is proportional to surface shear stress. In all boundary layers, u d*. Buckingham Pi theorem [282]: A mathematical theorem used in dimensional analysis that predicts the number of nondimensional groups that must be functionally related from a set of dimensional parameters that are thought to be functionally related. buffer layer [338, 580]: The part of a turbulent boundary layer, close to the wall, lying between the viscous and inertial sublayers. This thin layer is a transition from the frictiondominated layer adjacent to the wall where viscous stresses are large, to the inertial layer where turbulent stresses are large compared to viscous stresses. bulk modulus of elasticity [42]: See compressibility. buoyant force [89]: The net upward hydrostatic pressure force acting on an object submerged, or partially submerged, in a fluid. cavitation [40]: The formation of vapor bubbles in a liquid as a result of pressure going below the vapor pressure. center of pressure [79, 81]: The effective point of application of pressure distributed over a surface. This is the point where a counteracting force (equal to integrated pressure) must be placed for the net moment from pressure about that point to be zero. centripetal acceleration [250]: Acceleration associated with the change in the direction of the velocity (vector) of a material particle. closed system [14, 148]: See system. coefficient of compressibility [42, 55]: See compressibility. compressibility: The extent to which a fluid particle changes volume when subjected to either a change in pressure or a change in temperature. bulk modulus of elasticity [42, 55]: Synonymous with coefficient of compressibility. coefficient of compressibility [42, 55]: The ratio of pressure change to relative change in volume of a fluid particle. This coefficient quantifies compressibility in response to pressure change, an important effect in high Mach number flows. coefficient of volume expansion [44]: The ratio of relative density change to change in temperature of a fluid particle.
This coefficient quantifies compressibility in response to temperature change. computational fluid dynamics (CFD) [129, 296, 434]: The application of the conservation laws with boundary and initial conditions in mathematical discretized form to estimate field variables quantitatively on a discretized grid (or mesh) spanning part of the flow field. conservation laws [172]: The fundamental principles upon which all engineering analysis is based, whereby the material properties of mass, momentum, energy, and entropy can change only in balance with other physical properties involving forces, work, and heat transfer. These laws are predictive when written in mathematical form and appropriately combined with boundary conditions, initial conditions, and constitutive relationships. conservation of energy principle [201]: This is the first law of thermodynamics, a fundamental law of physics stating that the time rate of change of total energy of a fixed mass (system) is balanced by the net rate at which work is done on the mass and heat energy is transferred to the mass. Note: To mathematically convert the time derivative of mass, momentum, and energy of fluid mass in a system to that in a control volume, one applies the Reynolds transport theorem. conservation of mass principle [175]: A fundamental law of physics stating that a volume always containing the same atoms and molecules (system) must always contain the same mass. Thus the time rate of change of mass of a system is zero. This law of physics must be revised when matter moves at speeds approaching the speed of light so that mass and energy can be exchanged as per Einstein’s laws of relativity. conservation of momentum: This is Newton’s second law of motion, a fundamental law of physics stating that the time rate of change of momentum of a fixed mass (system) is balanced by the net sum of all forces applied to the mass. constitutive equations [426]: An empirical relationship between a physical variable in a conservation law of physics and other physical variables in the equation that are to be predicted. For example, the energy equation written for temperature includes the heat flux vector. It is known from experiments that heat flux for most common materials is accurately approximated as proportional to the gradient in temperature (this is called Fourier’s law). In Newton’s law written for a fluid particle, the viscous stress tensor (see stress) must be written as a function of velocity to solve the equation. The most common constitutive relationship for viscous stress is that for a Newtonian fluid. See also rheology. continuity equation [404]: Mathematical form of conservation of mass applied to a fluid particle in a flow. continuum [36, 122]: Treatment of matter as a continuous (without holes) distribution of finite mass differential volume
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elements. Each volume element must contain huge numbers of molecules so that the macroscopic effect of the molecules can be modeled without considering individual molecules. contour plot [138]: Also called an isocontour plot, this is a way of plotting data as lines of constant variable through a flow field. Streamlines, for example, may be identified as lines of constant stream function in two-dimensional incompressible steady flows. control mass [14]: See system. control volume [14, 122, 148]: A volume specified for analysis where flow enters and/or exits through some portion(s) of the volume surface. Also called an open system (see system). convective acceleration [126]: Synonymous with advective acceleration, this term must be added to the partial time derivative of velocity to properly quantify the acceleration of a fluid particle within an Eulerian frame of reference. For example, a fluid particle moving through a contraction in a steady flow speeds up as it moves, yet the time derivative is zero. The additional convective acceleration term required to quantify fluid acceleration (e.g., in Newton’s second law) is called the convective derivative. See also Eulerian description, Lagrangian description, material derivative, and steady flow. convective derivative: See material derivative and convective acceleration. creeping flow [313, 476, 574]: Fluid flow in which frictional forces dominate fluid accelerations to the point that the flow can be well modeled with the acceleration term in Newton’s second law set to zero. Such flows are characterized by Reynolds numbers that are small compared to 1 (Re 1). Since Reynolds number typically can be written as characteristic velocity times characteristic length divided by kinematic viscosity (VL /n), creeping flows are often slowmoving flows around very small objects (e.g., sedimentation of dust particles in air or motion of spermatozoa in water), or with very viscous fluids (e.g., glacier and tar flows). Also called Stokes flow. deformation rate [139]: See strain rate. derived dimensions [15]: See dimensions. deviatoric stress tensor [427]: Another term for viscous stress tensor. See stress. differential analysis [400]: Analysis at a point in the flow (as opposed to over a control volume). differential volume/area/length: A small volume dV, area dA, or length dx in the limit of the volume/area/length shrinking to a point. Derivatives are often produced in this limit. (Note that d is sometimes written as or d.) dimensional analysis [277]: A process of analysis based solely on the variables of relevance to the flow system under study, the dimensions of the variables, and dimensional homogeneity. After determining the other variables on which a variable of interest depends (e.g., drag on a car depends on
the speed and size of the car, fluid viscosity, fluid density, and surface roughness), one applies the principle of dimensional homogeneity with the Buckingham Pi theorem to relate an appropriately nondimensionalized variable of interest (e.g., drag) with the other variables appropriately nondimensionalized (e.g., Reynolds numbers, roughness ratio, and Mach number). dimensional homogeneity [273]: The requirement that summed terms must have the same dimensions (e.g., rV 2, pressure P, and shear stress txy are dimensionally . homogeneous while power, specific enthalpy h, and Pm are not). Dimensional homogeneity is the basis of dimensional analysis. dimensionality: The number of spatial coordinates in whose direction velocity components and/or other variables vary for a specified coordinate system. For example, fully developed flow in a tube is one-dimensional (1-D) in the radial direction r since the only nonzero velocity component (the axial, or x-, component) is constant in the x- and u-directions, but varies in the r-direction. Planar flows are two-dimensional (2-D). Flows over bluff bodies such as cars, airplanes, and buildings are three-dimensional (3-D). Spatial derivatives are nonzero only in the directions of dimensionality. dimensions [15, 270]: The required specification of a physical quantity beyond its numerical value. See also units. derived (or secondary) dimensions [15]: Combinations of fundamental dimensions. Examples of derived dimensions are: velocity (L/t), stress or pressure (F/L2 m/(Lt 2), energy or work (mL2/t 2 FL), density (m/L3), specific weight (F/L3), and specific gravity (unitless). fundamental (primary, basic) dimensions [15, 270]: Mass (m), length (L), time (t), temperature (T), electrical current (I), amount of light (C), and amount of matter (N) without reference to a specific system of units. Note that the force dimension is obtained through Newton’s law as F mL/t2 (thus, the mass dimension can be replaced with a force dimension by replacing m with Ft 2/L). drag coefficient [483, 565]: Nondimensional drag given by the drag force on an object nondimensionalized by dynamic pressure of the free-stream flow times frontal area of the object: CD 1
FD
2 rV
2
A
Note that at high Reynolds numbers (Re 1), CD is a normalized variable, whereas at Re 1, CD is nondimensional but is not normalized (see normalization). See also lift coefficient. drag force [46, 566]: The force on an object opposing the motion of the object. In a frame of reference moving with the object, this is the force on the object in the direction of flow. There are multiple components to drag force:
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friction drag [570]: The part of the drag on an object resulting from integrated surface shear stress in the direction of flow relative to the object. induced drag: The component of the drag force on a finite-span wing that is “induced” by lift and associated with the tip vortices that form at the tips of the wing and “downwash” behind the wing. pressure (or form) drag [570]: The part of the drag on an object resulting from integrated surface pressure in the direction of flow relative to the object. Larger pressure on the front of a moving bluff body (such as a car) relative to the rear results from massive flow separation and wake formation at the rear. dynamic pressure [189, 565]: When the Bernoulli equation in incompressible steady flow and/or the conservation of energy equation along a streamline are written in forms where each term in the equations has the dimensions force/area, dynamic pressure is the kinetic energy (per unit volume) term (i.e., 12 rV 2). dynamic similarity [278]: See similarity. dynamic viscosity [47]: See viscosity. dynamics [2]: When contrasted with statics the term refers to the application of Newton’s second law of motion to moving matter. When contrasted with kinematics the term refers to forces or accelerations through Newton’s law force balances. eddy viscosity [337]: See turbulence models. efficiency [184]: A ratio that describes levels of losses of useful power obtained from a device. Efficiency of 1 implies no losses in the particular function of the device for which a particular definition of efficiency is designed. For example, mechanical efficiency of a pump is defined as the ratio of useful mechanical power transferred to the flow by the pump to the mechanical energy, or shaft work, required to drive the pump. Pump-motor efficiency of a pump is defined as the ratio of useful mechanical power transferred to the flow over the electrical power required to drive the pump. Pump-motor efficiency, therefore, includes additional losses and is thus lower than mechanical pump efficiency. energy [41]: A state of matter described by the first law of thermodynamics that can be altered at the macroscopic level by work, and at the microscopic level through adjustments in thermal energy. flow energy [180]: Synonymous with flow work. The work associated with pressure acting on a flowing fluid. heat (transfer) [41]: The term “heat” is generally used synonymously with thermal energy. Heat transfer is the transfer of thermal energy from one physical location to another. internal energy [41]: Forms of energy arising from the microscopic motions of molecules and atoms, and from
the structure and motions of the subatomic particles comprising the atoms and molecules, within matter. kinetic energy [41]: Macroscopic (or mechanical) form of energy arising from the speed of matter relative to an inertial frame of reference. mechanical energy [182]: The nonthermal components of energy; examples include kinetic and potential energy. potential energy [41]: A mechanical form of energy that changes as a result of macroscopic displacement of matter relative to the gravitational vector. thermal energy [41]: Internal energy associated with microscopic motions of molecules and atoms. For singlephase systems, it is the energy represented by temperature. total energy [41]: Sum of all forms of energy. Total energy is the sum of kinetic, potential, and internal energies. Equivalently, total energy is the sum of mechanical and thermal energies. work energy [204]: The integral of force over the distance in which a mass is moved by the force. Work is energy associated with the movement of matter by a force. energy grade line [193]: See grade lines. English system [15]: See units. entry length [327]: The entry flow region in a pipe or duct flow where the wall boundary layers are thickening toward the center with axial distance x of the duct, so that axial derivatives are nonzero. As with the fully developed region, the hydrodynamic entry length involves growth of a velocity boundary layer, and the thermal entry length involves growth of a temperature boundary layer. Eulerian derivative [127]: See material derivative. Eulerian description [122]: In contrast with a Lagrangian description, an Eulerian analysis of fluid flow is developed from a frame of reference through which the fluid particles move. In this frame the acceleration of fluid particles is not simply the time derivative of fluid velocity, and must include another term, called convective acceleration, to describe the change in velocity of fluid particles as they move through a velocity field. Note that velocity fields are always defined in an Eulerian frame of reference. extensional strain rate [139]: See strain rate. extensive property [36, 150]: A fluid property that depends on total volume or total mass (e.g., total internal energy). See intensive property. field [124]: The representation of a flow variable as a function of Eulerian coordinates (x, y, z). For example, the velocity and acceleration fields are the fluid velocity and → → acceleration vectors (V , a ) as functions of position (x, y, z) in the Eulerian description at a specified time t. flow field [123]: The field of flow variables. Generally, this term refers to the velocity field, but it may also mean all field variables in a fluid flow.
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first law of thermodynamics [201]: See conservation laws, conservation of energy. flow separation [6, 569]: A phenomenon where a boundary layer adjacent to a surface is forced to leave, or “separate” from, the surface due to “adverse” pressure forces (i.e., increasing pressure) in the flow direction. Flow separation occurs in regions of high surface curvature, for example, at the rear of an automobile and other bluff bodies. flow work [205]: The work term in first law of thermodynamics applied to fluid flow associated with pressure forces on the flow. See energy, flow energy. fluid [2]: A material that when sheared deforms continuously in time during the period that shear forces are applied. By contrast, shear forces applied to a solid cause the material either to deform to a fixed static position (after which deformation stops), or cause the material to fracture. Consequently, whereas solid deformations are generally analyzed using strain and shear, fluid flows are analyzed using rates of strain and shear (see strain rate). fluid mechanics/dynamics [2]: The study and analysis of fluids through the macroscopic conservation laws of physics, i.e., conservation of mass, momentum (Newton’s second law), and energy (first law of thermodynamics), and the second law of thermodynamics. fluid particle/element [124]: A differential particle, or element, embedded in a fluid flow containing always the same atoms and molecules. Thus a fluid particle has fixed mass dm → and moves with the→flow with local flow velocity V , accel→ eration a particle DV /Dt and trajectory (xparticle(t), yparticle (t), tparticle (t)). See also material derivative, material particle, material position vector, and pathline. forced flow [11]: Flow resulting from an externally applied force. Examples include liquid flow through tubes driven by a pump and fan-driven airflow for cooling computer components. Natural flows, in contrast, result from internal buoyancy forces driven by temperature (i.e., density) variations within a fluid in the presence of a gravitational field. Examples include buoyant plumes around a human body or in the atmosphere. friction/frictional: See Newtonian fluid, viscosity, and viscous force. friction factor [331]: It can be shown from dimensional analysis and conservation of momentum applied to a steady fully developed pipe flow that the frictional contribution to the pressure drop along the pipe, nondimensionalized by flow dynamic pressure (12 rV 2avg), is proportional to the length-todiameter ratio (L /D) of the pipe. The proportionality factor f is called the friction factor. The friction factor is quantified from experiment (turbulent flow) and theory (laminar flow) in empirical relationships, and in the Moody chart, as a function of the Reynolds number and nondimensional roughness. Conservation of momentum shows that the friction factor is
proportional to the nondimensional wall shear stress (i.e., the skin friction). frictionless flow [191]: Mathematical treatments of fluid flows sometimes use conservation of momentum and energy equations without the frictional terms. Such mathematical treatments “assume” that the flow is “frictionless,” implying no viscous force (Newton’s second law), nor viscous dissipation (first law of thermodynamics). However, no real fluid flow of engineering interest can exist without viscous forces, dissipation, and/or head losses in regions of practical importance. The engineer should always identify the flow regions where frictional effects are concentrated. When developing models for prediction, the engineer should consider the role of these viscous regions in the prediction of variables of interest and should estimate levels of error in simplified treatments of the viscous regions. In high Reynolds number flows, frictional regions include boundary layers, wakes, jets, shear layers, and flow regions surrounding vortices. Froude number [274]: An order-of-magnitude estimate of the ratio of the inertial term in Newton’s law of motion to the gravity force term. The Froude number is an important nondimensional group in free-surface flows, as is generally the case in channels, rivers, surface flows, etc. fully developed [294, 325, 440]: Used by itself, the term is generally understood to imply hydrodynamically fully developed, a flow region where the velocity field is constant along a specified direction in the flow. In the fully developed region of pipe or duct flow, the velocity field is constant in the axial direction, x (i.e., it is independent of x), so that x-derivatives of velocity are zero in the fully developed region. There also exists the concept of “thermally fully developed” for the temperature field; however, unlike hydrodynamically fully developed regions where both the magnitude and shape of the velocity profile are constant in x, in thermally fully developed regions only the shape of the temperature profile is constant in x. See also entry length. fundamental dimensions [15, 270]: See dimensions. gage pressure [66]: Pressure (P) relative to atmospheric pressure (Patm). That is, Pgage P Patm. See also stress, pressure stress. Thus Pgage 0 or Pgage 0 is simply the pressure above or below atmospheric pressure. gas dynamics [2]: The study and analysis of gases and vapors through the macroscopic conservation laws of physics (see fluid mechanics/dynamics). geometric similarity: See similarity. grade lines [194]: Lines of head summations. energy grade line [195]: Line describing the sum of pressure head, velocity head, and elevation head. See head. hydraulic grade line [195]: Line describing the sum of pressure head and elevation head. See head.
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Hagen–Poiseuille flow: See Poiseuille flow. head [194]: A quantity (pressure, kinetic energy, etc.) expressed as an equivalent column height of a fluid. Conservation of energy for steady flow written for a control volume surrounding a central streamline with one inlet and one outlet, or shrunk to a streamline, can be written such that each term has the dimensions of length. Each of these terms is called a head term: elevation head [194]: The term in the head form of conservation of energy (see head) involving distance in the direction opposite to the gravitational vector relative to a predefined datum (z). head loss [330]: The term in the head form of conservation of energy (see head) that contains frictional losses and other irreversibilities. Without this term, the energy equation for streamlines becomes the Bernoulli equation in head form. pressure head [194]: The term in the head form of conservation of energy (see head) involving pressure (P/rg). velocity head [194]: The (kinetic energy) term in the head form of conservation of energy (see head) involving velocity (V 2/2g). heat [41]: See energy. hot-film anemometer [377]: Similar to a hot-wire anemometer except using a metallic film rather than a wire; used primarily for liquid flows. The measurement portion of a hot-film probe is generally larger and more rugged than that of a hot-wire probe. hot-wire anemometer [377]: A device used to measure a velocity component locally in a gas flow based on the relationship between the flow around a thin heated wire (the hot wire), temperature of the wire, and heating of the wire resulting from a current. See also hot-film anemometer. hydraulic grade line [193]: See grade lines. hydraulics [1]: The hydrodynamics of liquid and vapor flow in pipes, ducts, and open channels. Examples include water piping systems and ventilation systems. hydrodynamic entry length [325]: See entry length. hydrodynamically fully developed [325]: See fully developed. hydrodynamics [2]: The study and analysis of liquids through the macroscopic conservation laws of physics (see fluid mechanics/dynamics). The term is sometimes applied to incompressible vapor and gas flows, but when the fluid is air, the term aerodynamics is generally used instead. hydrostatic pressure [189]: The component of pressure variation in a fluid flow that would exist in the absence of flow as a result of gravitational body force. This term appears in the hydrostatic equation and in the Bernoulli equation. See also dynamic and static pressure. hypersonic [10]: An order of magnitude or more above the speed of sound (Mach number 1).
ideal fluid: See perfect fluid. ideal gas [38]: A gas at low enough density and/or high enough temperature that (a) density, pressure, and temperature are related by the ideal-gas equation of state, P rRT, and (b) specific internal energy and enthalpy are functions only of temperature. incompressible flow [10, 191, 406, 563]: A fluid flow where variations in density are sufficiently small to be negligible. Flows are generally incompressible either because the fluid is incompressible (liquids) or because the Mach number is low (roughly 0.3). induced drag [592]: See drag force. inertia/inertial: The acceleration term in Newton’s second law, or effects related to this term. Thus, a flow with higher inertia requires larger deceleration to be brought to rest. inertial sublayer [340]: A highly turbulent part of a turbulent boundary layer, close to the wall but just outside the viscous sublayer and buffer layer, where turbulent stresses are large compared to viscous stresses. intensive property [36, 150]: A fluid property that is independent of total volume or total mass (i.e., an extensive property per unit mass or sometimes per unit volume). internal energy [41]: See energy. inviscid (region of) flow [9, 327, 481]: Region of a fluid flow where viscous forces are sufficiently small relative to other forces (typically, pressure force) on fluid particles in that region of the flow to be neglected in Newton’s second law of motion to a good level of approximation (compare with viscous flow). See also frictionless flow. An inviscid region of flow is not necessarily irrotational. irrotational (region of) flow [144, 148, 325, 485, 579]: A region of a flow with negligible vorticity (i.e., fluid particle rotation). Also called potential flow. An irrotational region of flow is also inviscid. isocontour plot: See contour plot. jet: A friction-dominated region issuing from a tube or orifice and formed by surface boundary layers that have been swept behind by the mean velocity. Jets are characterized by high shear with the highest velocities in the center of the jet and lowest velocities at the edges. Frictional force, viscous stress, and vorticity are significant in jets. Kármán vortex street [133]: The two-dimensional alternating unsteady pattern of vortices that is commonly observed behind circular cylinders in a flow (e.g., the vortex street behind wires in the wind is responsible for the distinct tone sometimes heard). kinematic similarity [279]: See similarity. kinematic viscosity [48]: Fluid viscosity divided by density. kinematics [122]: In contrast with dynamics, the kinematic aspects of a fluid flow are those that do not directly involve
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Newton’s second law force balance. Kinematics refers to descriptions and mathematical derivations based only on conservation of mass (continuity) and definitions related to flow and deformation. kinetic energy [41]: See energy. kinetic energy correction factor [208]: Control volume analysis of the conservation of energy equation applied to tubes contains area integrals of kinetic energy flux. The integrals are often approximated as proportional to kinetic energy formed with area-averaged velocity, Vavg. The inaccuracy in this approximation can be significant, so a kinetic energy correction factor, a, multiplies the term to improve the approximation. The correction a depends on the shape of the velocity profile, is largest for laminar profiles (Poiseuille flow), and is closest to 1 in turbulent pipe flows at very high Reynolds numbers. Lagrangian derivative [127]: See material derivative. Lagrangian description [122]: In contrast with the Eulerian description, a Lagrangian analysis is developed from a frame of reference attached to moving material particles. For example, solid particle acceleration in the standard Newton’s → → second law form, F ma , is in a coordinate system that → moves with the particle so that acceleration a is given by the time derivative of particle velocity. This is the typical analytical approach used for analysis of the motion of solid objects. laminar flow [11, 323]: A stable well-ordered state of fluid flow in which all pairs of adjacent fluid particles move alongside one another forming laminates. A flow that is not laminar is either turbulent or transitional to turbulence, which occurs above a critical Reynolds number. laser Doppler velocimetry (LDV) [378]: Also called laser Doppler anemometry (LDA). A technique for measuring a velocity component locally in a flow based on the Doppler shift associated with the passage of small particles in the flow through the small target volume formed by the crossing of two laser beams. Unlike hot-wire and hot-film anemometry and like particle image velocimetry, there is no interference to the flow. lift coefficient [292, 565]: Nondimensional lift given by the lift force on a lifting object (such as an airfoil or wing) nondimensionalized by dynamic pressure of the free-stream flow times planform area of the object: CL 1
FL
2 rV
2
A
Note that at high Reynolds numbers (Re 1), CL is a normalized variable, whereas at Re 1, CL is nondimensional but is not normalized (see normalization). See also drag coefficient. lift force [566]: The net aerodynamic force on an object perpendicular to the motion of the object.
linear strain rate [139, 140]: Synonymous with extensional strain rate. See strain rate. losses [350]: Frictional head losses in pipe flows are separated into those losses in the fully developed pipe flow regions of a piping network, the major losses, plus head losses in other flow regions of the network, the minor losses. Minor loss regions include entry lengths, pipe couplings, bends, valves, etc. It is not unusual for minor losses to be larger than major losses. Mach number [10]: Nondimensional ratio of the characteristic speed of the flow to the speed of sound. Mach number characterizes the level of compressibility in response to pressure variations in the flow. major losses [347]: See losses. manometer [71]: A device that measures pressure based on hydrostatic pressure principles in liquids. material acceleration [127]: The acceleration of a fluid particle at the point (x, y, z) in a flow at time t.→This is given by the material derivative of fluid velocity: DV (x, y, z, t)/Dt. material derivative [127]: Synonymous terms are total derivative, substantial derivative, and particle derivative. These terms mean the time rate of change of fluid variables (temperature, velocity, etc.) moving with a fluid particle. Thus, the material derivative of temperature at a point (x, y, z) at time t is the time derivative of temperature attached to a moving fluid particle at the point (x, y, z) in the flow at the time t. In a Lagrangian frame of reference (i.e., a frame attached to the moving particle), particle temperature Tparticle depends only on time, so a time derivative is a total derivative dTparticle(t)/dt. In an Eulerian frame, the temperature field T(x, y, z, t) depends on both position (x, y, z) and time t, so the material derivative must include both a partial derivative in time and a convective derivative: dTparticle(t)/dt → → DT(x, y, z, t)/Dt T/t V T. See also field. material particle [124]: A differential particle, or element, that contains always the same atoms and molecules. Thus a material particle has fixed mass dm. In a fluid flow, this is the same as a fluid particle. material position vector [124]: A vector [xparticle(t), yparticle(t), zparticle(t)] that defines the location of a material particle as a function of time. Thus the material position vector in a fluid flow defines the trajectory of a fluid particle in time. mean: Synonymous with average. mechanical energy [180, 207]: See energy. mechanics [2]: The study and analysis of matter through the macroscopic conservation laws of physics (mass, momentum, energy, second law). minor losses [347]: See losses. mixing length [337]: See turbulence models. momentum: The momentum of a material particle (or fluid particle) is the mass of the material particle times its velocity.
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The momentum of a macroscopic volume of material particles is the integrated momentum per unit volume over the volume, where momentum per unit volume is the density of the material particle times its velocity. Note that momentum is a vector. momentum flux correction factor [236]: A correction factor added to correct for approximations made in the simplification of the area integrals for the momentum flux terms in the control volume form of conservation of momentum. Moody chart [341]: A commonly used plot of the friction factor as a function of the Reynolds number and roughness parameter for fully developed pipe flow. The chart is a combination of flow theory for laminar flow with a graphical representation of an empirical formula by Colebrook to a large set of experimental data for turbulent pipe flow of various values of “sandpaper” roughness. natural flow [11]: Contrast with forced flow. Navier–Stokes equation [429, 474]: Newton’s second law of fluid motion (or conservation of momentum) written for a fluid particle (the differential form) with the viscous stress tensor replaced by the constitutive relationship between stress and strain rate for Newtonian fluids. Thus the Navier–Stokes equation is simply Newton’s law written for Newtonian fluids. Newtonian fluid [47, 427]: When a fluid is subjected to a shear stress, the fluid continuously changes shape (deformation). If the fluid is Newtonian, the rate of deformation (i.e., strain rate) is proportional to the applied shear stress and the constant of proportionality is called viscosity. In general flows, the rate of deformation of a fluid particle is described mathematically by a strain rate tensor and the stress by a stress tensor. In flows of Newtonian fluids, the stress tensor is proportional to the strain rate tensor, and the constant of proportionality is called viscosity. Most common fluids (water, oil, gasoline, air, most gases and vapors) without particles or large molecules in suspension are Newtonian. Newton’s second law [230]: See conservation of momentum. nondimensionalization [272]: The process of making a dimensional variable dimensionless by dividing the variable by a scaling parameter (a single variable or a combination of variables) that has the same dimensions. For example, the surface pressure on a moving ball might be nondimensionalized by dividing it by rV2, where r is fluid density and V is freestream velocity. See also normalization. non-Newtonian fluid [427]: A non-Newtonian fluid is one that deforms at a rate that is not linearly proportional to the stress causing the deformation. Depending on the manner in which viscosity varies with strain rate, non-Newtonian fluids can be labeled shear thinning (viscosity decreases with increasing strain rate), shear thickening (viscosity increases
with increasing strain rate), and viscoelastic (when the shearing forces the fluid particles to return partially to an earlier shape). Suspensions and liquids with long-chain molecules are generally non-Newtonian. See also Newtonian fluid and viscosity. normal stress [3, 231]: See stress. normalization [272]: A particular nondimensionalization where the scaling parameter is chosen so that the nondimensionalized variable attains a maximum value that is of order 1 (say, within roughly 0.5 to 2). Normalization is more restrictive (and more difficult to do properly) than nondimensionalization. For example, P/(rV 2) discussed under nondimensionalization is also normalized pressure on a flying baseball (where Reynolds number Re 1), but is simply nondimensionalization of surface pressure on a small glass bead dropping slowly through honey (where Re 1). no-slip condition [6, 438]: The requirement that at the interface between a fluid and a solid surface, the fluid velocity and surface velocity are equal. Thus if the surface is fixed, the fluid must obey the boundary condition that fluid velocity 0 at the surface. one-dimensional [12]: See dimensionality. open system [14, 149]: Same as control volume. particle derivative [127, 129]: See material derivative. particle image velocimetry (PIV) [131, 380]: A technique for measuring a velocity component locally in a flow based on tracking the movement of small particles in the flow over a short time using pulsed lasers. Unlike hot-wire and hot-film anemometry and like laser Doppler velocimetry, there is no interference to the flow. pathline [130, 182]: A curve mapping the trajectory of a fluid particle as it travels through a flow over a period of time. Mathematically, this is the curve through the points mapped out by the material position vector [xparticle(t), yparticle(t), zparticle(t)] over a defined period of time. Thus, pathlines are formed over time, and each fluid particle has its own pathline. In a steady flow, fluid particles move along streamlines, so pathlines and streamlines coincide. In a nonsteady flow, however, pathlines and streamlines are generally very different. Contrast with streamline. perfect fluid: Also called an ideal fluid, the concept of a fictitious fluid that can flow in the absence of all frictional effects. There is no such thing as a perfect fluid, even as an approximation, so the engineer need not consider the concept further. periodic: An unsteady flow in which the flow oscillates about a steady mean. Pitot-static probe [190, 365]: A device used to measure fluid velocity through the application of the Bernoulli equation with simultaneous measurement of static and stagnation pressures. Also called a Pitot-Darcy probe.
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planar flow [419, 490]: A two-dimensional flow with two nonzero components of velocity in Cartesian coordinates that vary only in the two coordinate directions of the flow. Thus, all partial derivatives perpendicular to the plane of the flow are zero. See also axisymmetric flow and dimensionality. Poiseuille flow [316, 332]: Fully developed laminar flow in a pipe or duct. Also called Hagen–Poiseuille flow. The mathematical model relationships for Pouiseuille flow relating the flow rate and/or velocity profile to the pressure drop along the pipe/duct, fluid viscosity and geometry are sometimes referred to as Poiseuille’s law (although strictly not a “law” of mechanics). The velocity profile of all Poiseuille flows is parabolic, and the rate of axial pressure drop is constant. Poiseuille’s law [330]: See Poiseuille flow. potential energy [41]: See energy. potential flow [485]: Synonymous with irrotational flow. This is a region of a flow with negligible vorticity (i.e., fluid particle rotation). In such regions, a velocity potential function exists (thus the name). potential function [485]: If a region of a flow has zero vorticity (fluid particle spin), the velocity vector in that region can be written as the gradient of a scalar function called the velocity potential function, or simply the potential function. In practice, potential functions are often used to model flow regions where vorticity levels are small but not necessarily zero. power [202]: Work per unit time; time rate at which work is done. pressure [3, 66]: See stress. pressure force: As applicable to Newton’s second law, this is the force acting on a fluid particle that arises from spatial gradients in pressure within the flow. See also stress, pressure stress. pressure work: See flow work. primary dimension [15, 270]: See dimensions. profile plot [137]: A graphical representation of the spatial variation of a fluid property (temperature, pressure, strain rate, etc.) through a region of a fluid flow. A profile plot defines property variations in part of a field (e.g., a temperature profile might define the variation of temperature along a line within the temperature field). velocity profile [139]: The spatial variation in a velocity component or vector through a region of a fluid flow. For example, in a pipe flow the velocity profile generally defines the variation in axial velocity with radius across the pipe cross section, while a boundary layer velocity profile generally defines variation in axial velocity normal to the surface. The velocity profile is part of a velocity field. quasi-steady flow [475]: See steady flow.
Reynolds number [11, 279, 324]: An order-of-magnitude estimate of the ratio of the following two terms in Newton’s second law of motion over a region of the flow: the inertial (or acceleration) term over the viscous force term. Most but not all Reynolds numbers can be written as an appropriate characteristic velocity V times a characteristic length scale L consistent with the velocity V, divided by the kinematic viscosity n of the fluid: Re VL/n. The Reynolds number is arguably the most important nondimensional similarity parameter in fluid flow analysis since it gives a rough estimate of the importance of frictional force in the overall flow. Reynolds stress [337]: Velocity components (and other variables) in turbulent flows are separated into mean plus fluctuating components. When the equation for mean streamwise velocity component is derived from the Navier– Stokes equation, six new terms appear given by fluid density times the averaged product of two velocity components. Because these terms have the same units as stress (force/area), they are called turbulent stresses or Reynolds stresses (in memory of Osborne Reynolds who first quantified turbulent variables as mean + fluctuation). Just as viscous stresses can be written as a tensor (or matrix), we define a Reynolds stress tensor with Reynolds normal stress components and Reynolds shear stress components. Although Reynolds stresses are not true stresses, they have qualitatively similar effects as do viscous stresses, but as a result of the large chaotic vortical motions of turbulence rather than the microscopic molecular motions that underlie viscous stresses. Reynolds transport theorem [149]: The mathematical relationship between the time rate of change of a fluid property in a system (volume of fixed mass moving with the flow) and the time rate of change of a fluid property in a control volume (volume, usually fixed in space, with fluid mass moving across its surface). This finite volume expression is closely related to the material (time) derivative of a fluid property attached to a moving fluid particle. See also conservation laws. rheology [427]: The study and mathematical representation of the deformation of different fluids in response to surface forces, or stress. The mathematical relationships between stress and deformation rate (or strain rate) are called constitutive equations. The Newtonian relationship between stress and strain rate is the simplest example of a rheological constitutive equation. See also Newtonian and non-Newtonian fluid. rotation rate [139]: The angular velocity, or rate of spin, of a fluid particle (a vector, with units rad/s, given by 1/2 the curl of the velocity vector). See also vorticity. rotational flow [144, 146]: Synonymous with vortical flow, this term describes a flow field, or a region of a flow field, with significant levels of vorticity.
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926 FLUID MECHANICS
saturation pressure [39]: The pressure at which the phase of a simple compressible substance changes between liquid and vapor at fixed temperature. saturation temperature [39]: The temperature at which the phase of a simple compressible substance changes between liquid and vapor at fixed pressure. scaling parameter [273]: A single variable, or a combination of variables, that is chosen to nondimensionalize a variable of interest. See also nondimensionalization and normalization. schlieren technique [135]: An experimental technique to visualize flows based on the refraction of light from varying fluid density. The illuminance level in a schlieren image responds to the first spatial derivative of density. secondary dimensions [15]: See dimensions. shadowgraph technique [135]: An experimental technique to visualize flows based on the refraction of light from varying fluid density. The illuminance level in a shadowgraph image responds to the second spatial derivative of density. shear: Refers to gradients (derivatives) in velocity components in directions normal to the velocity component. shear force: See stress, shear stress. shear layer [147]: A quasi two-dimensional flow region with a high gradient in streamwise velocity component in the transverse flow direction. Shear layers are inherently viscous and vortical in nature. shear rate: The gradient in streamwise velocity in the direction perpendicular to the velocity. Thus, if streamwise (x) velocity u varies in y, the shear rate is du/dy. The term is applied to shear flows, where shear rate is twice the shear strain rate. See also strain rate. shear strain [139, 141]: See strain rate. shear stress [3, 231]: See stress, shear stress. shear thickening fluid [428]: See non-Newtonian fluid. shear thinning fluid [428]: See non-Newtonian fluid. SI system [15]: See units. similarity [277, 522]: The principle that allows one to quantitatively relate one flow to another when certain conditions are met. Geometric similarity, for example, must be true before one can hope for kinematic or dynamic similarity. The quantitative relationship that relates one flow to another is developed using a combination of dimensional analysis and data (generally, experimental, but also numerical or theoretical). dynamic similarity [278]: If two objects are geometrically and kinematically similar, then if the ratios of all forces (pressure, viscous stress, gravity force, etc.) between a point in the flow surrounding one object, and the same point scaled appropriately in the flow surrounding the other object, are all the same at all corresponding pairs of points, the flow is dynamically similar.
geometric similarity [277]: Two objects of different size are geometrically similar if they have the same geometrical shape (i.e., if all dimensions of one are a constant multiple of the corresponding dimensions of the other). kinematic similarity [277]: If two objects are geometrically similar, then if the ratios of all velocity components between a point in the flow surrounding one object, and the same point scaled appropriately in the flow surrounding the other object, are all the same at all corresponding pairs of points, the flow is kinematically similar. skin friction [523, 570]: Surface shear stress tw nondimensionalized by an appropriate dynamic pressure 1 2 2 rV . Also called the skin friction coefficient, Cf . solid: A material that when sheared either deforms to a fixed static position (after which deformation stops) or fractures. See also fluid. sonic [10]: At the speed of sound (Mach number 1). specific gravity [37]: Fluid density nondimensionalized by the density of liquid water at 4°C and atmospheric pressure (1 g/cm3 or 1000 kg/m3). Thus, specific gravity, SG r/rwater . specific weight [17, 237]: The weight of a fluid per unit volume, i.e., fluid density times acceleration due to gravity (specific weight, g rg). spin: See rotation rate and vorticity. stability [93]: A general term that refers to the tendency of a material particle or object (fluid or solid) to move away from or return when displaced slightly from its original position. neutrally stable [92]: See stability. When displaced slightly, the particle or object will remain in its displaced position. stable [92]: See stability. When displaced slightly, the particle or object will return to its original position. unstable [93]: See stability. When displaced slightly, the particle or object will continue to move from its original position. stagnation point [190]: A point in a fluid flow where the velocity goes to zero. For example, the point on the streamline that intersects the nose of a moving projectile is a stagnation point. stall [535, 570]: The phenomenon of massive flow separation from the surface of a wing when angle of attack exceeds a critical value, and consequent dramatic loss of lift and increase in drag. A plane in stall drops rapidly and must have its nose brought down to reestablish attached boundary layer flow and regenerate lift and reduce drag. static pressure [189]: Another term for pressure, used in context with the Bernoulli equation to distinguish it from dynamic pressure.
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927 GLOSSARY
statics [2]: The mechanical study and analysis of material that is fully at rest in a specific frame of reference. steady flow [11, 190]: A flow in which all fluid variables (velocity, pressure, density, temperature, etc.) at all fixed points in the flow are constant in time (but generally vary from place to place). Thus, in steady flows all partial derivatives in time are zero. Flows that are not precisely steady but that change sufficiently slowly in time to neglect time derivative terms with relatively little error are called quasi-steady. Stokes flow [478]: See creeping flow. strain [141]: See strain rate. strain rate [139]: Strain rate can also be called deformation rate. This is the rate at which a fluid particle deforms (i.e., changes shape) at a given position and time in a fluid flow. To fully quantify all possible changes in shape of a threedimensional fluid particle require six numbers. Mathematically, these are the six independent components of a secondrank symmetric strain rate tensor, generally written as a symmetric 3 3 matrix. Strain is time-integrated strain rate and describes deformation of a fluid particle after a period of time. See stress. extensional strain rate [139]: The components of strain rate that describe elongation or compression of a fluid particle in one of the three coordinate directions. These are the three diagonal elements of the strain rate tensor. The definition of extensional strain depends on one’s choice of coordinate axes. Also called linear strain rate. shear strain rate [139, 141]: The components of strain rate that describe deformation of a fluid particle in response to shear changing an angle between planes mutually perpendicular to the three coordinate axes. These are the off-diagonal elements of the strain rate tensor. The definition of shear strain depends on one’s choice of coordinate axes. volumetric strain rate [141]: Rate of change of volume of a fluid particle per unit volume. Also called bulk strain rate and rate of volumetric dilatation. streakline [132]: Used in flow visualization of fluid flows, this is a curve defined over time by the release of a marker (dye or smoke) from a fixed point in the flow. Contrast with pathline and streamline. In a steady flow, streamlines, pathlines, and streaklines all coincide. In a nonsteady flow, however, these sets of curves are each different from one another. stream function [309, 412]: The two velocity components in a two-dimensional steady incompressible flow can be defined in terms of a single two-dimensional function c that automatically satisfies conservation of mass (the continuity equation), reducing the solution of the two-component velocity field to the solution of this single stream function. This is done by writing the two velocity components as spatial derivatives of the stream function. A wonderful
property of the stream function is that (iso)contours of constant c define streamlines in the flow. streamline [129, 413, 493, 565]: A curve that is everywhere tangent to a velocity vector in a fluid velocity field at a fixed instant in time. Thus, the streamlines indicate the direction of the fluid motions at each point. In a steady flow, streamlines are constant in time and fluid particles move along streamlines. In a nonsteady flow the streamlines change with time and fluid particles do not move along streamlines. Contrast with pathline. streamtube [130]: A bundle of streamlines. A streamtube is usually envisioned as a surface formed by an infinite number of streamlines initiated within the flow on a circular circuit and tending to form a tubelike surface in some region of the flow. stress [3]: A component of a force distributed over an area is written as the integral of a stress over that area. Thus, stress is the force component dFi on a differential area element divided by the area of the element dAj (in the limit dAj → 0), where i and j indicate a coordinate direction x, y, or z. Stress sij dFi /dAj is therefore a force component per unit area in the i-direction on surface j. To obtain the surface force from stress, one integrates stress over the corresponding surface area. Mathematically, there are six independent components of a second-rank symmetric stress tensor, generally written as a symmetric 3 3 matrix. normal stress [3, 233]: A stress (force component per unit area) that acts perpendicular to the area. Therefore sxx, syy, and szz are normal stresses. The normal force over a surface is the net force from shear stress, given by integrating the shear stress over the surface area. The normal stresses are the diagonal elements of the stress tensor. pressure stress [3]: In a fluid at rest all stresses are normal stresses and all stresses act inward on a surface. At a fixed point, the three normal stresses are equal and the magnitude of these equal normal stresses is called pressure. Thus, in a static fluid sxx syy szz P, where P is pressure. In a moving fluid, stresses in addition to pressure are viscous stresses. A pressure force on a surface is the pressure stress integrated over the surface. The pressure force per unit volume on a fluid particle for Newton’s second law, however, is the negative of the gradient (spatial derivatives) of pressure at that point. Reynolds stress [339]: See Reynolds stress. shear stress [3, 233]: A stress (force component per unit area) that acts tangent to the area. Therefore, sxy, syx, sxz, szx, syz, and szy are shear stresses. The shear force over a surface is the net force from shear stress, given by integrating the shear stress over the surface area. The shear stresses are the off-diagonal elements of the stress tensor. turbulent stress [336, 337]: See Reynolds stress. viscous stress [429]: Flow creates stresses in the fluid that are in addition to hydrostatic pressure stresses. These
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928 FLUID MECHANICS
additional stresses are viscous since they arise from friction-induced fluid deformations within the flow. For example, sxx P txx, syy P tyy, and szz P tzz, where txx, tyy, and tzz are viscous normal stresses. All shear stresses result from friction in a flow and are therefore viscous stresses. A viscous force on a surface is a viscous stress integrated over the surface. The viscous force per unit volume on a fluid particle for Newton’s second law, however, is the divergence (spatial derivatives) of the viscous stress tensor at that point. stress tensor [233, 421]: See stress. subsonic [10]: Below the speed of sound (Mach number 1). substantial derivative [127]: See material derivative. supersonic [10]: Above the speed of sound (Mach number 1). surface tension [51]: The force per unit length at a liquid– vapor or liquid–liquid interface resulting from the imbalance in attractive forces among like liquid molecules at the interface. system [14, 148]: Usually when the word system is used by itself, closed system is implied, in contrast with a control volume or open system. closed system [14]: A volume specified for analysis that encloses always the same fluid particles. Therefore, no flow crosses any part of the volume’s surface and a closed system must move with the flow. Note that Newton’s law analysis of solid particles is generally a closed system analysis, sometimes referred to as a free body. open system [14]: A volume specified for analysis where flow crosses at least part of the volume’s surface. Also called a control volume. thermal energy [41]: See energy. three-dimensional [563]: See dimensionality. timeline [134]: Used for visualization of fluid flows, this is a curve defined at some instant in time by the release of a marker from a line in the flow at some earlier instant in time. The timeline, often used to approximate a velocity profile in a laboratory flow, is very different from streaklines, pathlines, and streamlines. tip vortex [592]: Vortex formed off each tip of an airplane wing as a byproduct of lift. Synonymous with trailing vortex. See also induced drag. total derivative [125, 127, 129]: See material derivative. total energy [41]: See energy. trailing vortex [595]: See tip vortex. trajectory: See pathline. transient period: A time-dependent period of flow evolution leading to a new equilibrium period that is generally, but not necessarily, steady. An example is the start-up period after a jet engine is switched on, leading to a steady (equilibrium) jet flow.
transitional flow [11, 323]: An unstable vortical fluid flow at a Reynolds number higher than a critical value that is large relative to 1, but is not sufficiently high that the flow has reached a fully turbulent flow state. Transitional flows often oscillate randomly between laminar and turbulent states. turbulence models [337]: Constitutive model relationships between Reynolds stresses and the mean velocity field in turbulent flows. Such model equations are necessary to solve the equation for mean velocity. A simple and widely used modeled form for the Reynolds stresses is to write them like the Newtonian relationship for viscous stresses, as proportional to the mean strain rate, with the proportionality being a turbulent viscosity or eddy viscosity. However, unlike Newtonian fluids, the eddy viscosity is a strong function of the flow itself, and the different ways in which eddy viscosity is modeled as a function of other calculated flow field variables constitute different eddy viscosity models. One traditional approach to modeling eddy viscosity is in terms of a mixing length, which is made proportional to a length set by the flow. turbulent flow [11, 323]: An unstable disordered state of vortical fluid flow that is inherently unsteady and that contains eddying motions over a wide range of sizes (or scales). Turbulent flows are always at Reynolds numbers above a critical value that is large relative to 1. Mixing is hugely enhanced, surface shear stresses are much higher, and head loss is greatly increased in turbulent flows as compared to corresponding laminar flows. turbulent stress [336, 337]: See Reynolds stress. turbulent viscosity [337]: See turbulence models. two-dimensional [562]: See dimensionality. units [15, 270]: A specific system to quantify numerically the dimensions of a physical quantity. The most common systems of units are SI (kg, N, m, s), English (lbm, lbf, ft, s), BGS (slug, lb, ft, s), and cgs (g, dyne, cm, s). See also dimensions. unsteady flow [11]: A flow in which at least one variable at a fixed point in the flow changes with time. Thus, in unsteady flows a partial derivative in time is nonzero for at least one point in the flow. vapor pressure [39]: The pressure below which a fluid, at a given temperature, will exist in the vapor state. See also cavitation and saturation pressure. velocity: A vector that quantifies the rate of change in position and the direction of motion of a material particle. velocity field [122]: See field. velocity profile [139]: See profile plots. viscoelastic fluid [427]: See non-Newtonian fluid. viscosity [46, 580]: See Newtonian fluid. Viscosity is a property of a fluid that quantifies the ratio of shear stress to rate of deformation (strain rate) of a fluid particle. (Therefore
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929 GLOSSARY
viscosity has the dimensions of stress/strain rate, or Ft/L2 m/Lt.) Qualitatively, viscosity quantifies the level by which a particular fluid resists deformation when subjected to shear stress (frictional resistance or friction). Viscosity is a measured property of a fluid and is a function of temperature. For Newtonian fluids, viscosity is independent of the rate of applied stress and strain rate. The viscous nature of nonNewtonian fluids is more difficult to quantify in part because viscosity varies with strain rate. The terms absolute viscosity, dynamic viscosity, and viscosity are synonymous. See also kinematic viscosity. viscous (regions of) flow [9]: Regions of a fluid flow where viscous forces are significant relative to other forces (typically, pressure force) on fluid particles in that region of the flow, and therefore cannot be neglected in Newton’s second law of motion (compare with inviscid flow). viscous (or frictional) force: As applicable to Newton’s second law, this is the force acting on a fluid particle that arises from spatial gradients in viscous (or frictional) stresses within the flow. The viscous force on a surface is the viscous stress integrated over the surface. See also stress, viscous stress. viscous stress tensor [427]: See stress. Also called the deviatoric stress tensor. viscous sublayer [340, 534, 580]: The part of a turbulent boundary layer adjacent to the surface that contains the
highest viscous stresses. The velocity gradient in this layer adjacent to the wall is exceptionally high. See also inertial layer and buffer layer. vortex: A local structure in a fluid flow characterized by a concentration of vorticity (i.e., fluid particle spin or rotation) in a tubular core with circular streamlines around the core axis. A tornado, hurricane, and bathtub vortex are common examples of vortices. Turbulent flow is filled with small vortices of various sizes, strengths, and orientations. vortical flow: Synonymous with rotational flow, this term describes a flow field, or a region of a flow field, with significant levels of vorticity. vorticity [144]: Twice the angular velocity, or rate of spin, of a fluid particle (a vector, with units rad/s, given by the curl of the velocity vector). See also rotation rate. wake [570]: The friction-dominated region behind a body formed by surface boundary layers that are swept to the rear by the free-stream velocity. Wakes are characterized by high shear with the lowest velocities in the center of the wake and highest velocities at the edges. Frictional force, viscous stress, and vorticity are significant in wakes. work [202]: See energy.
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CH CL , CL,x
Ca CD , CD,x Cd Cf , Cf,x
cp cv C C
c0
c
Bi Bo
B
bhp
b
Ar AR
A, Ac
a, a
→
a
Manning constant, m1/3/s; height from channel bottom to bottom of sluice gate, m Acceleration and its magnitude, m/s2 Area, m2; cross-sectional area, m2 Archimedes number Aspect ratio Width or other distance, m; intensive property in RTT analysis; turbomachinery blade width, m Brake horsepower, hp or kW Center of buoyancy; extensive property in RTT analysis Biot number Bond number Specific heat for incompressible substance, kJ/kg ! K; speed of sound, m/s; speed of light in a vacuum, m/s; chord length of an airfoil, m Wave speed, m/s Constant-pressure specific heat, kJ/kg ! K Constant-volume specific heat, kJ/kg ! K Dimension of the amount of light Bernoulli constant, m2/s2 or m/t2 ! L, depending on the form of Bernoulli equation; Chezy coefficient, m1/2/s; circumference, m Cavitation number Drag coefficient; local drag coefficient Discharge coefficient Fanning friction factor or skin friction coefficient; local skin friction coefficient Head coefficient Lift coefficient; local lift coefficient F, F FB FD Ff FL
→
f, fx
E . E, E Ec EGL Es Eu f
DAB Dh Dp e → → e r, e u
CNPSH CP Cp CP CQ CS CV Cwd D or d
Suction head coefficient Center of pressure Pressure coefficient Power coefficient Capacity coefficient Control surface Control volume Weir discharge coefficient Diameter, m (d typically for a smaller diameter than D) Species diffusion coefficient, m2/s Hydraulic diameter, m Particle diameter, m Specific total energy, kJ/kg Unit vector in r- and u-direction, respectively Voltage, V Total energy, kJ; and rate of energy, kJ/s Eckert number Energy grade line, m Specific energy in open-channel flows, m Euler number Frequency, cycles/s; Blasius boundary layer dependent similarity variable Darcy friction factor; and local Darcy friction factor Force and its magnitude, N Magnitude of buoyancy force, N Magnitude of drag force, N Magnitude of drag force due to friction, N Magnitude of lift force, N
N O M E N C L AT U R E
→
Ixx j
I I
i
→
Hgross i
HGL
→
H, H
hfg hL H
Gr h
GM
G
. g
g, g
FT
Fo Fr
Fourier number Froude number Magnitude of tension force, N Gravitational acceleration and its magnitude, m/s2 Heat generation rate per unit volume, W/m3 Center of gravity Metacentric height, m Grashof number Specific enthalpy, kJ/kg; height, m; head, m; convective heat transfer coefficient, W/m2 ! K Latent heat of vaporization, kJ/kg Head loss, m Boundary layer shape factor; height, m; net head of a pump or turbine, m; total energy of a liquid in open-channel flow, expressed as a head, m; weir head, m Moment of momentum and its magnitude, N!m!s Hydraulic grade line, m Gross head acting on a turbine, m Index of intervals in a CFD grid (typically in x-direction) Unit vector in x-direction Dimension of electric current Moment of inertia, N ! m ! s2; current, A; turbulence intensity Second moment of inertia, m4 Reduction in Buckingham Pi theorem; index of intervals in a CFD grid (typically in y-direction)
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→
NSp
NPSH
NP
N N
n
→
. n, n
Ma n
Lh Lw m . m, m M → M, M
L L Le Lc
!
KE KL Kn
K
ke
k
→
Ja k
j
Unit vector in y-direction Jakob number Specific heat ratio; expected number of "s in buckingham Pi theorem; thermal conductivity, W/m ! K; turbulent kinetic energy per unit mass, m2/s2; index of intervals in a CFD grid (typically in z-direction) Unit vector in z-direction Specific kinetic energy, kJ/kg Doublet strength, m3/s Kinetic energy, kJ Minor loss coefficient Knudsen number Length or distance, m; turbulent length scale, m Dimension of length Length or distance, m Lewis number Chord length of an airfoil, m; characteristic length, m Hydrodynamic entry length, m Weir length, m Dimension of mass Mass, kg; and mass flow rate, kg/s Molar mass, kg/kmol Moment of force and its magnitude, N!m Mach number Number of parameters in Buckingham Pi theorem; Manning coefficient Number of rotations; and rate of rotation, rpm Unit normal vector Dimension of the amount of matter Number of moles, mol or kmol; number of blades in a turbomachine Power number Net positive suction head, m Pump specific speed SG
Sc Sf
Sc
S0
Ri Ru s
Rh
Ra Re
R
QEAS → r, r
. Q, Q
PE Pe Pgage Pm Pr Psat or Pv Pvac Pw q . q
Nu p pe P, P#
NSt
Heat flux (rate of heat transfer per unit area), W/m2 Total heat transfer, kJ; and rate of heat transfer, W or kW Equiangle skewness in a CFD grid Moment arm and its magnitude, m; radial coordinate, m; radius, m Gas constant, kJ/kg ! K; radius, m; electrical resistance, $ Rayleigh number Reynolds number Hydraulic radius, m Richardson number Universal gas constant, kJ/kmol ! K Submeerged distance along the plane of a plate, m; distance along a surface or streamline, m; specific entropy, kJ/kg ! K; fringe spacing in LDV, m; turbomachinery blade spacing, m Slope of the bottom of a channel in openchannel flow Schmidt number Critical slope in open-channel flow Friction slope in open-channel flow Specific gravity
Turbine specific speed Nusselt number Wetted perimeter, m Specific potential energy, kJ/kg Pressure and modified pressure, N/m2 or Pa Potential energy, kJ Peclet number Gage pressure, N/m2 or Pa Mechanical pressure, N/m2 or Pa Prandtl number Saturation pressure or vapor pressure, kPa Vacuum pressure, N/m2 or Pa Weir height, m Heat transfer per unit mass, kJ/kg
x
→
We x
W . W, W
w
V0
v . V, V → V, V
v
U
uz
uu
ur
u*
Sh SP St Stk t t T T → T, T u
Sherwood number Property at a stagnation point Stanton number; Strouhal number Stokes number Dimension of time Time, s Dimension of temperature Temperature, °C or K Torque and its magnitude, N ! m Specific internal energy, kJ/kg; Cartesian velocity component in x-direction, m/s Friction velocity in turbulent boundary layer, m/s Cylindrical velocity component in r-direction, m/s Cylindrical velocity component in u-direction, m/s Cylindrical velocity component in z-direction, m/s Internal energy, kJ; x-component of velocity outside a boundary layer (parallel to the wall), m/s Cartesian velocity component in y-direction, m/s Specific volume, m3/kg Volume, m3; and volume flow rate, m3/s Velocity and its magnitude (speed), m/s; average velocity, m/s Uniform-flow velocity in open-channel flow, m/s Work per unit mass, kJ/kg; Cartesian velocity component in z-direction, m/s; width, m Weight, N; width, m Work transfer, kJ; and rate of work (power), W or kW Weber number Cartesian coordinate (usually to the right), m Position vector, m
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Cartesian coordinate (usually up or into the page), m; depth of liquid in openchannel flow, m Normal depth in open-channel flow, m Cartesian coordinate (usually up), m
→
"
m n n(Ma)
l
k
eij & f gs ' h
d* e
d
b
a, a
correction factor; thermal diffusivity, m2/s; isothermal compressibility, kPa%1 or atm%1 Angular acceleration and its magnitude, s%2 Coefficient of volume expansion, K%1; momentum-flux correction factor; angle; diameter ratio in obstruction flowmeters; oblique shock angle; turbomachinery blade angle Boundary layer thickness, m; distance between streamlines, m; angle; small change in a quantity Boundary layer displacement thickness, m Mean surface roughness, m; turbulent dissipation rate, m2/s3 Strain rate tensor, s%1 Dissipation function, kg/m ! s3 Angle; velocity potential function, m2/s Specific weight, N/m3 Circulation or vortex strength, m2/s Efficiency; Blasius boundary layer independent similarity variable Bulk modulus of compressibility, kPa or atm; log law constant in turbulent boundary layer Mean free path length, m; wavelength, m; second coefficient of viscosity, kg/m ! s Viscosity, kg/m ! s; Mach angle Kinematic viscosity m2/s Prandtl–Meyer function for expansion waves, degrees or rad Nondimensional parameter in dimensional analysis
Greek Letters a Angle; angle of attack; kinetic energy
z
yn
y
→
Angle or angular coordinate; boundary layer momentum thickness, m; pitch angle of a turbomachinery blade; turning or deflection angle of oblique shock Density, kg/m3 Normal stress, N/m2 Stress tensor, N/m2 Surface tension, N/m Shear stress, N/m2 Viscous stress tensor (also called shear stress tensor), N/m2 Specific Reynolds stress tensor, m2/s2 Angular velocity vector and its magnitude, rad/s; angular frequency, rad/s Stream function, m2/s Vorticity vector and its magnitued, s%1
cr CL CS CV e eff f H lam L m
C c
abs atm avg b
0
Stagnation property; property at the origin or at a reference point Absolute Atmospheric Average quantity Property of the back or exit of a nozzle, e.g., back pressure Pb Acting at the centroid Pertaining to a cross section Critical property Pertaining to the centerline Pertaining to a control surface Pertaining to a control volume Property at an exit; extracted portion Effective property Property of a fluid, usually of a liquid Acting horizontally Property of a laminar flow Portion lost by irreversibilities Property of a model
Subscripts ( Property of the far field
z, z
c →
v, v
tij, turbulent
r s sij ss t tij
u
Maximum value Mechanical property Minimum value Normal component Acting at the center of pressure Property of a prototype; property of a particle; property of a piston Resultant Relative (moving frame of reference) Rectangular property Property of the rotor leading edge Property of the rotor trailing edge Acting on a surface Property of a solid Saturation property; property of a satellite Property of the stator leading edge Property of the stator trailing edge Submerged portion Pertaining to a system Tangential component Trianglular property Property of a turbulent flow Useful portion Acting vertically Property of a vapor Vacuum Property at a wall
Averaged quantity Quantity per unit time; time derivative # (prime) Fluctuating quantity; derivative of a variable; modified variable * Nondimensional property; sonic property ) Law of the wall variable in turbulent boundary layer → (over arrow) Vector quantity
(overbar) ! (overdot)
Superscripts _
w
v vac
u V
tri turb
t
sat sl st sub sys
S s
r rec rl rt
R
P p
n
max mech min
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Errata Sheet for Fluid Mechanics: Fundamentals and Applications – Çengel and Cimbala Latest update: 02-08-2007 This is a list of errors (and enhancements) in the textbook. If you find any additional errors in the book, or have suggestions for improvement, please contact John M. Cimbala at 814-863-2739 or [email protected] to report it. [By way of acknowledgment, the person (other than the authors) who first reports an error is listed in brackets, unless requested otherwise.] Note: First check the copyright page to see which printing you have. At the time of this writing, there have been two printings, and the third printing is coming shortly. The errors and enhancements are listed according to printing number in reverse chronological order to save you time. For example, all the errors and enhancements listed for the first printing have been fixed in the second printing; if you have the second printing, ignore all the corrections listed for the first printing. For each printing, we categorize the changes as major errors, minor errors, or enhancements: • Major errors are important and significant (e.g., incorrect equations or numerical values) – these must be changed. • Minor errors are spelling or typo errors and other minor changes – these may be skipped without impacting understanding of the material. • Enhancements are changes that clarify something and/or help you to understand the material better (e.g., improvements to a figure or wording changes) – these may be skipped since they are not really errors, but are useful changes that enhance understanding of the material.
Corrections in the third printing (September 2006) – “3 4 5 6 7 8 9 0 DOW/DOW 0 9 8 7 6” on the copyright page. These errors were found after the deadline for the 3rd printing, and will be corrected in the 4th printing (date to be determined). Make these changes only if you have the 1st, 2nd, or 3rd printing of the book. Major Errors in the Third Printing • Pg. 47, Eq. 2-26: Change “tan β ” to “tan (dβ )”. [Jason DeGraw] • Pg. 56, left column, first line of text above References: Change “droplet or bubble is” to “droplet or soap bubble, respectively, is”. • Pg. 58, Prob. 2-10 table: Change the title of the first column from “z, km” to “r, km”. • Pg. 94, Fig. 3-46a: Move the labels and arrows for B up, and move the labels and arrows for G down so that B is above G as in Fig. 3-44(a). [Junfeng Zhang] • Pg. 189, first line: Change “Eq. 5-44” to “Eq. 5-43”. [Jason DeGraw] • Pg. 190, third line above the new section: Change “free-stream value” to “free-stream velocity”. [Jason DeGraw] • Pg. 257, Fig. 6-38: Reverse the direction of angular velocity, omega. [Milivoje Kostic] • Pg. 261, Prob. 6-20: At the end, add “Assume the jet stream is perpendicular to the incoming liquid flow line”. • Pg. 340, lines 2 and 6 of the last paragraph before new section: Change “laminar sublayer” to “viscous sublayer” (two occurrences). [Rustom Bhiladvala] • Pg. 342, line 10 of the last bullet item: Change “laminar sublayer” to “viscous sublayer”. [Rustom Bhiladvala] • Pg. 344, Fig. 8-30: The labels are mixed up: “2 in” is the dimension and “0.2 ft3/s water” is the flow rate. [Jason DeGraw] • Pg. 371, line 3 of figure caption: Change “close-up view” to “cutaway view”. [Jason DeGraw] • Pg. 371, lines 5 to 6 of figure caption: Remove “looking down the axis with flow into the page,”. [Jason DeGraw] • Pg. 375, Fig. 8-69: Turn the pipe in the middle figure to the same orientation (same flow direction) as the left figure. • Pg. 512, Fig. 10-77b: Move the dimension arrow for δ (x) up a bit to the top of the plate. • Pg. 543, Eq. (1) of Example 10-15: Add a third (missing) measurement, namely “V = 10.0 m/s”. [Jose Sinibaldi] • Pg. 571, Fig. 11-17: Add an x axis label, namely, “Re”. [Jason DeGraw] • Pg. 807, Prob. 14-36, 3rd line from end of problem: Change “m(Lpm)2” to “m/(Lpm)2”. [Alex Schwartz] Minor Errors in the Third Printing • Pg. 17, line 2 above Eq. 1-2: Change “from Newton’s” to “from an equation based on Newton’s”. • Pg. 42, Eq. 2-9: Change the font for the v in cv to the volume font, not the velocity font, i.e., change to . • Pg. 59, Prob. 2-35: Add at the end of the problem statement, “For simplicity, assume β = constant = β at 40oC” [Rustom Bhiladvala] • Pg. 149, Fig. 4-52: Add a pointer from the word “System” pointing to the dashed line around the left spray can. [Jason DeGraw] • Pg. 190, 6 lines before the new section: Change “parallel the z-axis” to “parallel to the z-axis”. [Mohamed Elsari] • Pg. 242, second and third equations on the page: Remove the arrow from V – it is a scalar (two occurrences). • Pg. 249, sixth line on blackboard of Fig. 6-28: Add an arrow over the M – it is a vector. [Jason DeGraw] • Pg. 251, Fig. 6-31: Remove the extraneous plus sign on the upper right part of the figure. • Pg. 251, Eqs. 6-43 and 6-44: Change the font for V to that of volume, not velocity (two occurrences). [Jason DeGraw] • Pg. 252, Eq. 6-46 and last equation in Fig. 6-33: Add an arrow over the H – it is a vector. [Jason DeGraw]
• • • • • • • • •
• • •
Pg. 271, Fig. 7-5: Change “WATCHOUT” to “WATCH OUT” (two words, not one) on the sign. [Jason DeGraw] Pg. 291, Fig. 3: Make the f italic, i.e., change “f” to “f ”. Pg. 345, last equation in the cluster of four equations near the bottom: Change “20 =” to “20 m =”. Pg. 346, last equation in the cluster of four equations near the bottom: Change “20 =” to “20 m =”. Pg. 352, lines 1 and 2: Change “convert” to “converts”. Pg. 352, line above Eq. 8-60: Change “the case of sudden expansion is” to “the case of a sudden expansion is”. Pg. 368, line 5: Change “head loss is small.” to “head loss is smaller.”. [Jason DeGraw] Pg. 484, line below Eq. 1: Change “with radius as” to “with radius (across streamlines) as”. Pg. 599, Reference 9: Change “J. Happel” to “J. Happel and H. Brenner”. Also change “Hydrocarbons” to “Hydrodynamics”. [Rustom Bhiladvala] Pg. 603, Fig. P11-35: The two horizontal cups are oriented wrong. Turn 90o. [Jason DeGraw] Pg. 616, 5th line below Eq. (b): Change the reference from “Çengel, 2002” to “Çengel and Boles, 2002”. Pg. 687, last line of first bullet item near the bottom of page: Change “flow energy” to “potential energy”. Pg. 697, line 3 of blue highlighted statement and line 4 of the figure caption for Fig. 13-21: Change “cross-section” to “cross-sectional area” (two occurrences on this page). Pg. 699, lines 2 to 3 below Eq. 13-46: Change “cross section” to “cross-sectional area”. Pg. 713, line after the equation for hL, near the middle: Change “dissipation ratio is” to “dissipation ratio are”. Pg. 724, lines 3 to 4 of paragraph 2, left column: Change “cross section” to “cross-sectional area”.
•
Back inside cover, entry for ζ , ζ : Change “magnitued” to “magnitude”.
• • • •
G
Enhancements to the Third Printing G • Pg. 239, line after Eq. 6-29: Change “on the body. This” to “on the body. We remind ourselves that V in E. 6-29 is the fluid velocity relative to an inertial reference frame. This”. [Krishna Pillai] • Pg. 266, Prob. 6-65, line 5: Change “10 m/s” to “10 m/s relative to the skater”. • Pg. 264, Prob. 6-48E, last sentence: Change “the electric power” to “the maximum possible electric power”. • Pg. 338, Fig. 8-24: Add this sentence at the end of the caption: “Note that u(r) in the turbulent case is the time-averaged velocity component in the axial direction (the overbar on u has been dropped for simplicity)”. [Rustom Bhiladvala] • Pg. 352, 2 lines above Eq. 8-60: Change “the conservation of mass, momentum, and energy equations” to “the equations of mass, momentum, and energy balance”. • Pg. 484, line above Eq. 3: Change “coordinates reduces to” to “coordinates (see Eq. 9-62b without the viscous terms) reduces to”. [Cliff Moses]
Corrections in the second printing (July 2005) – “2 3 4 5 6 7 8 9 0 DOW/DOW 0 9 8 7 6 5” on the copyright page. These were corrected in the 3rd printing (expected Aug. 2006). Make these changes only if you have the 1st or 2nd printing of the book. Major Errors in the Second Printing • Pg. 63, Problem 2-76E, line 2: Change “1408” to “140o”. [Woonjean Park] • Pg. 68, 2nd line below Eq. 3-5: Change “xz-plane” to “yz-plane”. [Woonjean Park] • Pg. 107, Problem 3-47, line 2: Change “by a piston” to “by a 30-cm-diameter piston”. • Pg. 108, Answer to problem 3-48: Change “5.0” to “1.34”. [Woonjean Park] • Pg. 110, Problem 3-66, first line: Change “A 4-m-high” to “A 5-m-high”. [Tao Xing] • Pg. 110, Figure P3-68E: Change the dimension “5 ft” to “8 ft” for consistency with the solutions manual. [Roscoe Ward] • Pg. 111, Figure P3-82: Change the dimension “10 cm” to “9.5 cm” for consistency with the problem statement. [Tao Xing] • Pg. 113, Problem 3-101, last line: Change “1030 kg/m” to “1030 kg/m3”. [Russ LaBarca] G G G • Pg. 231, Fig. 6-5: Add the subscript “body” to the label dF on the left. I.e., change “ dF ” to “ dFbody ”. • • • • • • • • • •
Pg. 261, Answers to Prob. 6-27: Change “(a) 250 N, (b) 1.25 kW” to “(a) -167 N, (b) 833 W”. [Ted Thiede] Pg. 263, Prob. 6-37, 2nd line: Change “30 m/s” to “30 m/s relative to the ground”. Pg. 310, Answer to Problem 7-33: Change “V/V 2∞” to “V 2/V∞2”. [Woonjean Park] Pg. 328, Eq. 8-14: Change “1” to “r2” in the numerator of the first term on the right side of the equation. Pg. 356, 3 lines below Eq. 8-62: Change “nearly 1” to “about 1.05”. Pg. 388, Answer to Problem 8-45: Change “93.8 m” to “10.8 m”. [Woonjean Park] Pg. 431, top middle term on right side of Eq. 9-63: Change “ ∂ r ” to “ ∂r ”, i.e., the r should not be a subscript. Pg. 466, Fig. P9-99: Change the lower vertical dimension from “h2” to “h1”. [Russ LaBarca] G G Pg. 468, answer to Problem 9-111: Change “ VP ” to “ ∇P ”. Pg. 524, Figure 10-102: All the streamlines, a pointer line, and a dashed line are missing, as shown below. [Eric Paterson] V
Outer flow streamline
U(x) = V
δ*(x)
y
δ(x)
• • • • • • •
x Boundary layer Pg. 554, Prob. 10-48: In all three equations, φ should not be a subscript, i.e., change “∂φ” to “∂φ” (3 places). Pg. 555, 3rd line of Problem 10-54: Change “(−ay + cx)” to “(−ay + c)”. [Ted Thiede] Pg. 603, Answer given for Problem 11-41: Change “186” to “135”. [Russ LaBarca] Pg. 620, Eq. 12-16: The division line is missing between the numerator dA and the denominator A. [Eric Paterson] Pg. 709, 2nd equation on the page – the equation for “yc =”: In the denominator, change “(12 m2)” to “(12 m2)2”. Pg. 709, 2nd equation on the page and also on the first line of Discussion: Change the answer from “2.2 m” to “2.65 m”. Pg. 927, entry for normal stress (subset of stress): Change “shear” to “normal” in two places.
Minor Errors in the Second Printing • Pg. ix, end of Chapter 4: Move the Summary entry up, just above the entry for the Application Spotlight. [Woonjean Park] • Pg. xi, end of Chapter 10: Move the Summary and References entries up, just above the entry for the Application Spotlight. [Woonjean Park] • Pg. xi, end of Chapter 11: Move the Summary and References entries up, just above the entry for the Application Spotlight. [Woonjean Park] • Pg. 8, line 5 of second paragraph: Change “Osborn” to “Osborne”. • Pg. 9, lines 1 and 5: Change “Theodore” to “Theodor” (two places). [Note – both spellings are commonly used, but “Theodore” is an Americanized version; the Hungarian name does not include the “e” at the end] • Pg. 17, figure 1-28: Change “=” to “≈” in three places. • Pg. 42, Eq. 2-10 (twice), 3 lines below Eq. 2-10, Eq. 2-11, and 1 line below Eq. 2-11 (5 occurrences total): Change subscript “ave” to “avg”. • Pg. 57, line 3 of second paragraph: Change “and ranging” to “and ranging” [“a” not italicized]. [Woonjean Park] • Pg. 73, Figure 3-15: Extend the line for the lower limit of dimension h2 to the right such that it nearly touches the interface between the oil (light blue) and the mercury (dark blue). Also extend the lower limit of dimension h3. [Woonjean Park] • Pg. 85, first two equations in the Analysis (2 places): Change the subscript “ave” to “avg”. [Woonjean Park] • Pg. 87, first two equations in the Analysis (2 places): Change the subscript “ave” to “avg”. [Woonjean Park] • Pg. 89, Eq. 3-33 (2 places): Change the subscript “ave” to “avg”. [Woonjean Park]
• • • • • • • • • •
• • • •
Pg. 102, second-to-last equation in the right column: Change the subscript “ave” to “avg”. [Woonjean Park] Pg. 109, Figure P3-52: Two corrections – (1) color the fluid in the line from the bottom of the tank to point A grey, the same color as the glycerin. (2) remove the grey piston-like shading at the top (it is open to the air). [Jim Brasseur] Pg. 112, Answer to Prob. 3-88: Change the subscript “ave” to “avg”. [Woonjean Park] Pg. 116, Problem 3-126E, line 2: Change “1026E” to “126E”. [Woonjean Park] Pg. 153, 3 places; Pg. 154, 11 places; Pg. 155, 3 places; Pg. 167, 1 place: Change “Leibnitz” to “Leibniz”. [Both spellings appear in the literature, but the one without the “t” is the proper spelling] [Clement Kleinstreuer] Nomenclature on back inside cover, page 2, left column, line 4 under the entry for k: Change “buckingham” to “Buckingham”. [Wonnam Lee] Nomenclature on inside back cover, page 2, middle column, line 11 from bottom under the entry for s: Change “submeerged” to “submerged”. [Wonnam Lee] Pg. 449, 4th line of Assumptions: Change “constant-pressure” to “constant pressure”, i.e., remove the hyphen. Pg. 687, two lines below Eq. 13-18: Change “of constant” to “of rectangular cross section and of constant”. [Jason DeGraw] Pg. 931 and throughout the Index: Remove entries for variables and units. In particular, remove the entry for A, atm, Bar, bhp, Btu, C, cd, Fr, ft, g, GM, Havailable, Hpump, u, Hrequired, J, J/m2, K, kg, kgf, kgf/cm2, Kilogram-force per square centimeter, kJ, kPa, lbf, lbf/ft, lbf/in2, lbm, m, m2/s, Ma, MPa, N, N⋅m, N⋅m/m2, N/m, N/m2, NPSH, NPSHrequired, P*, psi, R, Rgas, s, and Wwater horsepower. Pg. 938: Remove all three entries for “ER”, i.e., “ER clutch, 60”, “ER effect, 60”, and “ER fluids, 60”. Pg. 948, entry for “Pressure prism”: Change the page numbers from “82” to “80,82”. Pg. 954, second column: Remove entry for “TT”. Pg. 956, third column: Remove all entries for the letter “X”.
Enhancements to the Second Printing • Pg. 32, line 3 of Problem 1-18: Remove the word “upward”, since the direction of the acceleration is not given. • Pg. 37, line 2 above Eq. 2-3: Change “is called specific weight and” to “is called specific weight, or weight density, and”. • Pg. 40, line 14: Change “higher temperatures boils at higher pressures” to “higher pressure boils at higher temperature”. [Woonjean Park] • Pg. 45, end of line 2 and beginning of line 3: Change “density difference, which is proportional” to “density difference, which is in turn proportional”. [Woonjean Park] • Pg. 53, lines 1, 3, 5 from top, lines 1, 2, and 3 below Eq. 2-37, in Fig. 2-22b, and in the caption of Fig. 2-22b (8 occurrences total): Change “bubble” to “soap bubble” for clarity – we are not talking about air bubbles in water. • Pg. 53, label to the left of Eq. 2-37: Change “Bubble” to “Soap bubble”. • Pg. 53, line 4 below Eq. 2-37: Change “bubble consisting” to “existence”. • Pg. 53, first line of first new paragraph below Eq. 2-37: Change “droplet (or bubble)” to “droplet of liquid in a gas (or a bubble of gas in a liquid)”. • Pg. 66, 5th line above Eq. 3-1: Change “Pressures below atmospheric pressure are called” to “Pgage can be positive or negative, but pressures below atmospheric pressure are sometimes called”. [Jim Brasseur] • Pg. 66, 2nd line above Eq. 3-1: Delete “all positive quantities and are”. [Jim Brasseur] G • Pg. 68, Fig. 3-4: Add an arrow in the white space pointing down, and label it g as in Fig. 3-48. G • Pg. 69, Fig. 3-6: Add an arrow in the white space pointing down, an label it g as in Fig. 3-48. • Pg. 69, Fig. 3-6: Remove the gray shading in the figure. • Pg. 80, Figures 3-25 and 3-26 and Eq. 3-19 (3 places): Change the subscript “ave” to “avg”. [Woonjean Park] • Pg. 80, Fig. 3-25: Change the label on the right from “ ” to “Pressure prism”. •
Pg. 80, Fig. 3-25: Change the equation label on the bottom right from “
•
Park] Pg. 82, line 4: Change the in-line equation from “
• • • • • • •
” to “ FR = ∫ PdA ”. [Woonjean
” to “ FR = ∫ PdA ”.
Pg. 83, Caption of Fig. 3-29, line 2: Change “form a volume” to “form a pressure prism”. Pg. 108, Figure P3-48: Extend the dimension for 22 cm vertically such that it is at the same elevation as the interface between the air (white) and fluid 1 (dark blue). [Woonjean Park] Pg. 113, Problem 3-102, last two lines: Change “pressure and its location” to “pressure difference and the location of the maximum pressure”. [Tao Xing] Pg. 118, Figure P3-134: Add small horizontal lines at the top and bottom of the dimension arrow for Δh. [Woonjean Park] Pg. 119, Problem 3-144, line 5: Change “the same.” to “the same, and the fluids meet at the axis of rotation.”. Pg. 119, Figure P3-144: Several errors: (1) Remove the gray and light blue shaded areas. (2) Move the short vertical line in the bottom tube (the interface between the left and right fluids) to the left such that it is coincident with the axis of rotation. (3) color the fluid on the left side gray (everything up to the short vertical line at its revised location). [Woonjean Park] Pg. 122, Right in the middle: Change “parcels of fluid” to “fluid particles”.
• • • • • • • • • • •
Pg. 125, 4th line below Eq. 4-8: Change “time t, since the” to “time t. Why? Because the”. G G Pg. 177, line 6 of the section “Moving or…”: Change “ VCV ” to “ VCS ” (two occurrences). Pg. 177, line 7 of the section “Moving or…”: Change “control volume” to “control surface”. G Pg. 187, Fig. 5-23: Add an arrow in the white space pointing down, and label it g as in Fig. 3-48. Pg. 194, given info for Example 5-5: Change to: “Water is flowing from a garden hose (Fig. 5-38). A child places his thumb to cover most of the hose outlet, causing a thin jet of high-speed water to emerge. The pressure in the hose just upstream of his thumb is 400 kPa. If the hose is held upward, what is the maximum height that the jet could achieve?” [Jim Brasseur] Pg. 194, last line: Remove Assumption 2. Then, on Pg. 195, renumber Assumption 3 to 2, 4 to 3, and 5 to 4. [Jim Brasseur] Pg. 195, line 4 above the first equation on the page: Change “V1” to “V12 m min and thus the net force acting on it is greater than the weight of the plate, and the difference between the jet impulse and the weight will accelerate the plate upwards. Therefore, Newton’s 2nd law F = ma = mdV/dt in this case can be expressed as m (V J − V ) − W = m p a
→
ρA(V J − V ) 2 − m p g = m p
mpg
ρA
V
m p dV
0
ρ A(VJ − V ) − m p g 2
=∫
t t =0
dt
→
t=∫
mpg m − ρA ρA
FRz
Guide rails
W = mp g
dV dt
Separating the variables and integrating from t = 0 when V = 0 to t = t when V = V gives the desired integral,
∫
→ V=
V
m p dV
0
ρ A(VJ − V )2 − m p g
Discussion This integral can be performed with the help of integral tables. But the relation obtained will be implicit in V.
mp
. m
Nozzle
6-51 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 6 Momentum Analysis of Flow Systems 6-71 Solution Water enters a centrifugal pump axially at a specified rate and velocity, and leaves at an angle from the axial direction. The force acting on the shaft in the axial direction is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The forces acting on the piping system in the horizontal direction are negligible. 3 The atmospheric pressure is disregarded since it acts on all surfaces. 4 Water flow is nearly uniform at the outlet and thus the momentum-flux correction factor can be taken to be unity, β ≅ 1. Properties
We take the density of water to be 1000 kg/m3.
Analysis From conservation of mass we have m 1 = m 2 = m , and thus V1 = V2 and Ac1V1 = Ac 2V 2 . Noting that the discharge area is half the inlet area, the discharge velocity is twice the inlet velocity. That is, A Ac1V 2 = c1 V1 = 2V1 = 2(5 m/s ) = 10 m/s Ac 2
We take the pump as the control volume, and the inlet direction of flow as the positive direction of x axis. The linear momentum equation in this case in the x direction reduces to G G G F= βm V − β m V → − FRx = m V 2 cosθ − m V1 →
∑
∑ out
∑
z x 60 °
m V
FRx
n
FRx = m (V1 − V 2 cosθ )
in
where the mass flow rate it m = ρV = (1000 kg/m 3 )(0.20 m 3 /s) = 200 kg/s Substituting the known quantities, the reaction force is determined to be (note that cos60° = 0.5) ⎛ 1N FRx = (200 kg/s)[(5 m/s) - (10 m/s)cos60]⎜⎜ 2 ⎝ 1 kg ⋅ m/s
⎞ ⎟=0 ⎟ ⎠
Discussion Note that at this angle of discharge, the bearing is not subjected to any horizontal loading. Therefore, the loading in the system can be controlled by adjusting the discharge angle.
6-72 Solution Water enters the impeller of a turbine through its outer edge of diameter D with velocity V making an angle α with the radial direction at a mass flow rate of m , and leaves the impeller in the radial direction. The maximum power sinα . that can be generated is to be shown to be Wshaft = π nmDV Assumptions negligible.
α
1 The flow is steady in the mean. 2 Irreversible losses are
Analysis We take the impeller region as the control volume. The tangential velocity components at the inlet and the outlet are V1,t = 0 and
G V
r2 =D/2
G V
V 2,t = V sin α .
Normal velocity components as well pressure acting on the inner and outer circumferential areas pass through the shaft center, and thus they do not contribute to torque. Only the tangential velocity components contribute to torque, and the application of the angular momentum equation gives Tshaft = m (r2V 2,t − r1V1,t ) = m r2V 2,t − 0 = m D(V sin α ) / 2
ω
Tshaft Impeller region
The angular velocity of the propeller is ω = 2πn . Then the shaft power becomes W shaft = ωTshaft = 2πn m D (V sin α ) / 2 Simplifying, the maximum power generated becomes W shaft = πn m DV sin α which is the desired relation. Discussion
The actual power is less than this due to irreversible losses that are not taken into account in our analysis.
6-52 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 6 Momentum Analysis of Flow Systems 6-73 Solution A two-armed sprinkler is used to water a garden. For specified flow rate and discharge angles, the rates of rotation of the sprinkler head are to be determined.
θ r = 0.45 m •
θ
Assumptions 1 The flow is uniform and cyclically steady (i.e., steady from a frame of reference rotating with the sprinkler head). 2 The water is discharged to the atmosphere, and thus the gage pressure at the nozzle outlet is zero. 3 Frictional effects and air drag of rotating components are neglected. 4 The nozzle diameter is small compared to the moment arm, and thus we use average values of radius and velocity at the outlet.
We take the density of water to be 1000 kg/m3 = 1 kg/L.
Properties
Analysis We take the disk that encloses the sprinkler arms as the control volume, which is a stationary control volume. The conservation of mass equation for this steady flow system is m 1 = m 2 = m . Noting that the two nozzles are identical, we have m nozzle = m / 2 or Vnozzle = Vtotal / 2 since the density of water is constant. The average jet outlet velocity relative to the nozzle is V jet =
⎛ 1m3 ⎞ Vnozzle 60 L/s ⎜ ⎟ = 95.49 m/s = 2 A jet 2[π (0.02 m) / 4] ⎜⎝ 1000 L ⎟⎠
The angular momentum equation can be expressed as
∑ M = ∑ rm V − ∑ rm V . Noting that there are no external out
in
moments acting, the angular momentum equation about the axis of rotation becomes 0 = −2rm nozzleV r cos θ
→
Vr = 0
→
V jet,t −V nozzle = 0
Noting that the tangential component of jet velocity is V jet, t = V jet cos θ , we have V nozzle = V jet cos θ = (95.49 m/s)cosθ
Also noting that V nozzle = ωr = 2πn r , and angular speed and the rate of rotation of sprinkler head become 1) θ = 0°: ω =
Vnozzle ( 95.49 m/s ) cos0 = = 212 rad/s 0.45 m r
and
n =
ω 212 rad/s ⎛ 60 s ⎞ = ⎜ ⎟ = 2026 rpm ≅ 2030 rpm 2π 2π ⎝ 1 min ⎠
2) θ = 30°: ω =
Vnozzle ( 95.49 m/s ) cos30° ω 184 rad/s ⎛ 60 s ⎞ = = 184 rad/s and n = = ⎜ ⎟ = 1755 rpm ≅ 1760 rpm 0.45 m 2π 2π r ⎝ 1 min ⎠
3) θ = 60°: ω =
Vnozzle ( 95.49 m/s ) cos60° ω 106 rad/s ⎛ 60 s ⎞ = = 106 rad/s and n = = ⎜ ⎟ = 1013 rpm ≅ 1010 rpm 0.45 m 2π 2π r ⎝ 1 min ⎠
Discussion Final results are given to three significant digits, as usual. The rate of rotation in reality will be lower because of frictional effects and air drag.
6-53 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 6 Momentum Analysis of Flow Systems
6-74 Solution The previous problem is reconsidered. The effect of discharge angle θ on the rate of rotation n as θ varies from 0 to 90° in increments of 10° is to be investigated. Analysis
The EES Equations window is printed below, followed by the tabulated and plotted results.
D=0.02 "m" r=0.45 "m" n_nozzle=2 "number of nozzles" Ac=pi*D^2/4 V_jet=V_dot/Ac/n_nozzle V_nozzle=V_jet*cos(theta) V_dot=0.060 "m3/s" omega=V_nozzle/r n_dot=omega*60/(2*pi)
Angle, θ° 0 10 20 30 40 50 60 70 80 90
n rpm 2026 1996 1904 1755 1552 1303 1013 693 352 0
ω rad/s 212 209 199 184 163 136 106 73 37 0
Vnozzle , m/s 95.5 94.0 89.7 82.7 73.2 61.4 47.7 32.7 16.6 0.0
2250
n, rpm
1800
1350
900
450
0 0
10
20
30
40
θ, °
50
60
70
80
90
Discussion The maximum rpm occurs when θ = 0o, as expected, since this represents purely tangential outflow. When θ o = 90 , the rpm drops to zero, as also expected, since the outflow is purely radial and therefore there is no torque to spin the sprinkler.
6-54 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 6 Momentum Analysis of Flow Systems 6-75 Solution A stationary water tank placed on wheels on a frictionless surface is propelled by a water jet that leaves the tank through a smooth hole. Relations are to be developed for the acceleration, the velocity, and the distance traveled by the tank as a function of time as water discharges. Assumptions 1 The orifice has a smooth entrance, and thus the frictional losses are negligible. 2 The flow is steady, incompressible, and irrotational (so that the Bernoulli equation is applicable). 3 The surface under the wheeled tank is level and frictionless. 4 The water jet is discharged horizontally and rearward. 5 The mass of the tank and wheel assembly is negligible compared to the mass of water in the tank. 4 Jet flow is nearly uniform and thus the momentum-flux correction factor can be taken to be unity, β ≅ 1. Analysis (a) We take point 1 at the free surface of the tank, and point 2 at the outlet of the hole, which is also taken to be the reference level (z2 = 0) so that the water height above the hole at any time is z. Noting that the fluid velocity at the free surface is very low (V1 ≅ 0), it is open to the atmosphere (P1 = Patm), and water discharges into the atmosphere (and thus P2 = Patm), the Bernoulli equation simplifies to V12
V22
P P1 + + z1 = 2 + + z2 ρg 2 g ρg 2 g
→
z=
VJ2 2g
+0
→
1
z x
D
VJ = 2 gz
D0
VJ
2
The discharge rate of water from the tank through the hole is m = ρAV J = ρ
πD02 4
VJ = ρ
πD02 4
2 gz
The momentum equation for steady one-dimensional flow is
G
G
G
∑ F = ∑ βm V − ∑ βm V . Applying it to the water tank, out
in
the horizontal force that acts on the tank is determined to be F = m V e − 0 = m V J = ρ
πD02
2 gz = ρgz
πD02
4 2 The acceleration of the water tank is determined from Newton’s 2nd law of motion F = ma where m is the mass of water in the tank, m = ρV tank = ρ (πD 2 / 4 ) z ,
a=
2 F ρ gz (π D0 / 2 ) = m ρ z (π D 2 / 4 )
a = 2g
→
D02 D2
Note that the acceleration of the tank is constant. (b) Noting that a = dV/dt and thus dV = adt and acceleration a is constant, the velocity is expressed as D02 t D2 (c) Noting that V = dx/dt and thus dx = Vdt, the distance traveled by the water tank is determined by integration to be V = at
→
dx = Vdt
→
V = 2g
dx = 2 g
D02 tdt D2
→
x=g
D02 2 t D2
since x = 0 at t = 0. Discussion In reality, the flow rate discharge velocity and thus the force acting on the water tank will be less because of the frictional losses at the hole. But these losses can be accounted for by incorporating a discharge coefficient.
Design and Essay Problems 6-76 Solution
Students’ essays and designs should be unique and will differ from each other.
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Chapter 7 Dimensional Analysis and Modeling
Solutions Manual for Fluid Mechanics: Fundamentals and Applications by Çengel & Cimbala
CHAPTER 7 DIMENSIONAL ANALYSIS AND MODELING
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Chapter 7 Dimensional Analysis and Modeling
Dimensions and Units, Primary Dimensions
7-1C Solution
We are to explain the difference between a dimension and a unit, and give examples.
Analysis A dimension is a measure of a physical quantity (without numerical values), while a unit is a way to assign a number to that dimension. Examples are numerous – length and meter, temperature and oC, weight and lbf, mass and kg, time and second, power and watt,… Discussion units.
7-2C Solution
When performing dimensional analysis, it is important to recognize the difference between dimensions and
We are to append the given table with other parameters and their primary dimensions.
Analysis Students’ tables will differ, but they should add entries such as angular velocity, kinematic viscosity, work, energy, power, specific heat, thermal conductivity, torque or moment, stress, etc. Discussion This problem should be assigned as an ongoing homework problem throughout the semester, and then collected towards the end. Individual instructors can determine how many entries to be required in the table.
7-3C Solution
We are to list the seven primary dimensions and explain their significance.
Analysis The seven primary dimensions are mass, length, time, temperature, electrical current, amount of light, and amount of matter. Their significance is that all other dimensions can be formed by combinations of these seven primary dimensions. Discussion One of the first steps in a dimensional analysis is to write down the primary dimensions of every variable or parameter that is important in the problem.
7-4 Solution
We are to write the primary dimensions of the universal ideal gas constant.
Analysis
From the given equation,
Primary dimensions of the universal ideal gas constant: ⎧ m ⎫ × L3 ⎪ ⎧ pressure × volume ⎫ ⎪⎪ t 2 L ⎪ {Ru } = ⎨ ⎬=⎨ ⎬= ⎩ mol × temperature ⎭ ⎪ N × T ⎪ ⎩⎪ ⎭⎪
⎧ mL2 ⎫ ⎨ 2 ⎬ ⎩ t TN ⎭
Or, in exponent form, {Ru} = {m1 L2 t-2 T-1 N-1}. Discussion the result.
The standard value of Ru is 8314.3 J/kmol⋅K. You can verify that these units agree with the dimensions of
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Chapter 7 Dimensional Analysis and Modeling 7-5 Solution
We are to write the primary dimensions of atomic weight.
Analysis
By definition, atomic weight is mass per mol,
{M } = ⎧⎨
mass ⎫ ⎧ m ⎫ ⎬= ⎨ ⎬ ⎩ mol ⎭ ⎩ N ⎭
Primary dimensions of atomic weight:
(1)
Or, in exponent form, {M} = {m1 N-1}. Discussion In terms of primary dimensions, atomic mass is not dimensionless, although many authors treat it as such. Note that mass and amount of matter are defined as two separate primary dimensions.
7-6 Solution We are to write the primary dimensions of the universal ideal gas constant in the alternate system where force replaces mass as a primary dimension. Analysis
From Newton’s second law, force equals mass times acceleration. Thus, mass is written in terms of force as
Primary dimensions of mass in the alternate system: 2 force ⎫ ⎧ F ⎫ ⎧ Ft ⎫ {mass} = ⎧⎨ ⎬=⎨ 2⎬=⎨ ⎬ ⎩ acceleration ⎭ ⎩ L/t ⎭ ⎩ L ⎭
(1)
We substitute Eq. 1 into the results of Problem 7-4, Primary dimensions of the universal ideal gas constant:
⎧ Ft 2 2 ⎫ L ⎪ ⎧ mL ⎫ ⎪ FL {Ru } = ⎨ 2 ⎬ = ⎨⎪ L 2 ⎬⎪ = ⎧⎨ ⎫⎬ ⎩ Nt T ⎭ ⎪ Nt T ⎪ ⎩ TN ⎭ ⎪⎩ ⎪⎭ 2
(2)
Or, in exponent form, {Ru} = {F1 L1 T-1 N-1}. Discussion Eq. 2.
The standard value of Ru is 8314.3 J/kmol⋅K. You can verify that these units agree with the dimensions of
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Chapter 7 Dimensional Analysis and Modeling 7-7 Solution We are to write the primary dimensions of the specific ideal gas constant, and verify are result by comparing to the standard SI units of Rair. Analysis We can approach this problem two ways. If we have already worked through Problem 7-4, we can use our results. Namely,
Primary dimensions of specific ideal gas constant:
{R } gas
⎧ mL2 ⎫ ⎧ R ⎫ ⎪⎪ 2 ⎪⎪ ⎧ L2 ⎫ = ⎨ u ⎬ = ⎨ Nt T ⎬ = ⎨ 2 ⎬ ⎩ M ⎭ ⎪ m ⎪ ⎩t T⎭ ⎩⎪ N ⎭⎪
(1)
Or, in exponent form, {Rgas} = {L2 t-2 T-1}. Alternatively, we can use either form of the ideal gas law, Primary dimensions of specific ideal gas constant:
{R } gas
⎧ m ⎫ × L3 ⎪ ⎧ 2 ⎫ ⎧ pressure × volume ⎫ ⎪⎪ t 2 L L ⎪ =⎨ ⎬=⎨ ⎬= ⎨ 2 ⎬ t × × mass temperature m T ⎩ ⎭ ⎪ ⎪ ⎩ T⎭ ⎪⎩ ⎭⎪
(2)
For air, Rair = 287.0 J/kg⋅K. We transform these units into primary dimensions, Primary dimensions of the specific ideal gas constant for air: ⎧ mL2 ⎫ ⎧ J ⎫ ⎪⎪ t 2 ⎪⎪ {Rair } = ⎨287.0 ⎬=⎨ ⎬= kg×K ⎭ ⎪ m × T ⎪ ⎩ ⎩⎪ ⎭⎪
⎧ L2 ⎫ ⎨ 2 ⎬ ⎩t T⎭
(3)
Equation 3 agrees with Eq. 1 and Eq. 2, increasing our confidence that we have performed the algebra correctly. Discussion
Notice that numbers, like the value 287.0 in Eq. 3 have no influence on the dimensions.
7-8 Solution
We are to write the primary dimensions of torque and list its units.
Analysis
Torque is the product of a length and a force,
G
Primary dimensions of torque:
⎫ ⎧ L ⎫ ⎬ = ⎨m ⎬ {M } = ⎧⎨⎩length × mass length time ⎭ ⎩ t ⎭ 2
2
2
(1)
G Or, in exponent form, { M } = {m1 L2 t-2}. The common units of torque are newton-meter (SI) and inch-pound (English). In primary units, however, we write the primary SI units according to Eq. 1,
Primary SI units:
Units of torque = kg ⋅ m 2 /s 2
and in primary English units, Primary English units:
Units of torque = lbm ⋅ ft 2 /s 2
Discussion Since torque is the product of a force and a length, it has the same dimensions as energy. Primary units are not required for dimensional analysis, but are often useful for unit conversions and for verification of proper units when solving a problem.
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Chapter 7 Dimensional Analysis and Modeling 7-9 Solution
We are to determine the primary dimensions of each variable.
Analysis (a) Energy is force times length (the same dimensions as work),
Primary dimensions of energy: 2 mass × length ⎫ ⎧ mL ⎫ length × = ⎬ ⎨ 2 ⎬ time2 ⎩ ⎭ ⎩ t ⎭
{E} = {force × length} = ⎧⎨
(1)
Or, in exponent form, {E} = {m1 L2 t-2}. (b) Specific energy is energy per unit mass, Primary dimensions of specific energy: energy ⎫ ⎧ mass × length 2 1 ⎫ ⎧ L2 ⎫ × {e} = ⎧⎨ ⎬=⎨ ⎬=⎨ ⎬ mass ⎭ ⎩ t 2 ⎭ time 2 ⎩ mass ⎭ ⎩
(2)
Or, in exponent form, {e} = { L2 t-2}. (c) Power is the rate of change of energy, i.e. energy per unit time, Primary dimensions of power: 2 1 ⎫ ⎧ mL2 ⎫ ⎧ energy ⎫ ⎧ mass × length W = ⎨ × ⎬=⎨ ⎬=⎨ ⎬ 2 time ⎭ ⎩ t 3 ⎭ time ⎩ time ⎭ ⎩
{ }
(3)
Or, in exponent form, { W } = {m1 L2 t-3}. Discussion
In dimensional analysis it is important to distinguish between energy, specific energy, and power.
7-10 Solution
We are to determine the primary dimensions of electrical voltage.
Analysis
From the hint,
Primary dimensions of voltage: ⎧ mass × length 2 power ⎫ ⎪⎪ time3 {voltage} = ⎨⎧ ⎬=⎨ current ⎩ current ⎭ ⎪ ⎪⎩
⎫ 2 ⎪⎪ ⎧ mL ⎫ = ⎬ ⎨ 3 ⎬ ⎪ ⎩ t I ⎭ ⎪⎭
(1)
Or, in exponent form, {E} = {m1 L2 t-3 I-1}. Discussion dimensions.
We see that all dimensions, even those of electrical properties, can be expressed in terms of primary
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Chapter 7 Dimensional Analysis and Modeling 7-11 Solution Analysis current,
We are to write the primary dimensions of electrical resistance. From Ohm’s law, we see that resistance has the dimensions of voltage difference divided by electrical
Primary dimensions of resistance: ⎧ mass × length 2 ⎫ ⎪ ⎪ ⎧ mL2 ⎫ 3 ΔE {R} = ⎧⎨ ⎫⎬ = ⎪⎨ time × current ⎪⎬ = ⎨ 3 2 ⎬ current ⎩ I ⎭ ⎪ ⎪ ⎩t I ⎭ ⎩⎪ ⎭⎪
Or, in exponent form, {R} = {m1 L2 t-3 I-2}, where we have also used the result of the previous problem. Discussion
All dimensions can be written in terms of primary dimensions.
7-12 Solution
We are to determine the primary dimensions of each variable.
Analysis (a) Acceleration is the rate of change of velocity,
Primary dimensions of acceleration:
{a} = ⎧⎨
velocity ⎫ ⎧ length 1 ⎫ ⎧L ⎫ × ⎬=⎨ ⎬=⎨ 2⎬ ⎩ time ⎭ ⎩ time time ⎭ ⎩ t ⎭
(1)
Or, in exponent form, {a} = { L1 t-2}. (b) Angular velocity is the rate of change of angle, Primary dimensions of angular velocity:
{ω} = ⎧⎨
angle ⎫ ⎧ 1 ⎫ ⎧ 1 ⎫ ⎬=⎨ ⎬=⎨ ⎬ ⎩ time ⎭ ⎩ time ⎭ ⎩ t ⎭
(2)
Or, in exponent form, {ω} = { t-1}. (c) Angular acceleration is the rate of change of angular velocity, Primary dimensions of angular acceleration: angular velocity ⎫ ⎧ 1 1 ⎫ ⎧1⎫ × {α = ω } = ⎧⎨ ⎬=⎨ ⎬=⎨ 2⎬ time ⎩ ⎭ ⎩ time time ⎭ ⎩ t ⎭
(3)
Or, in exponent form, {α} = { t-2}. Discussion In Part (b) we note that the unit of angle is radian, which is a dimensionless unit. Therefore the dimensions of angle are unity.
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Chapter 7 Dimensional Analysis and Modeling 7-13 Solution
We are to write the primary dimensions of angular momentum and list its units.
Analysis
Angular momentum is the product of length, mass, and velocity,
Primary dimensions of angular momentum: G length ⎫ ⎧ mL2 ⎫ ⎧ (1) H = ⎨length × mass × ⎬=⎨ ⎬ t time ⎩ ⎭ ⎩ ⎭ G Or, in exponent form, { H } = {m1 L2 t-1}. We write the primary SI units according to Eq. 1,
{ }
Units of angular momentum =
Primary SI units:
kg ⋅ m 2 s
and in primary English units, Units of angular momentum =
Primary English units:
lbm ⋅ ft 2 s
Discussion Primary units are not required for dimensional analysis, but are often useful for unit conversions and for verification of proper units when solving a problem.
7-14 Solution
We are to determine the primary dimensions of each variable.
Analysis (a) Specific heat is energy per unit mass per unit temperature,
Primary dimensions of specific heat at constant pressure:
{c } p
⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ t 2 ⎪⎪ ⎧ L2 ⎫ energy =⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩ mass × temperature ⎭ ⎪ m × T ⎪ ⎩ t T ⎭ ⎩⎪ ⎭⎪
(1)
Or, in exponent form, {cp} = { L2 t-2 T-1}. (b) Specific weight is density times gravitational acceleration,
⎧ mass length ⎫ ⎧ m ⎫ =⎨ 2 2⎬ Primary dimensions of specific weight: { ρ g} = ⎨ 2 ⎬ ⎩ volume time ⎭ ⎩ L t ⎭
(2)
Or, in exponent form, {ρg} = {m1 L-2 t-2}. (c) Specific enthalpy has dimensions of energy per unit mass, ⎧ mL2 ⎧ energy ⎫ ⎪⎪ t 2 Primary dimensions of specific enthalpy: {h} = ⎨ ⎬=⎨ ⎩ mass ⎭ ⎪ m ⎪⎩
⎫ 2 ⎪⎪ ⎧ L ⎫ ⎬=⎨ 2 ⎬ ⎪ ⎩t ⎭ ⎪⎭
(3)
Or, in exponent form, {h} = { L2 t-2}. Discussion As a check, from our study of thermodynamics we know that dh = cpdT for an ideal gas. Thus, the dimensions of dh must equal the dimensions of cp times the dimensions of dT. Comparing Eqs. 1 and 3 above, we see that this is indeed the case.
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Chapter 7 Dimensional Analysis and Modeling 7-15 Solution
We are to determine the primary dimensions of thermal conductivity.
Analysis The primary dimensions of Q conduction are energy/time, and the primary dimensions of dT/dx are temperature/length. From the given equation,
Primary dimensions of thermal conductivity:
⎧ ⎫ ⎧ mL2 ⎫ energy ⎪⎪ ⎪⎪ ⎪⎪ t 3 ⎪⎪ ⎧ mL ⎫ time {k } = ⎨ ⎬=⎨ ⎬=⎨ 3 ⎬ ⎪ length 2 × temperature ⎪ ⎪ L × T ⎪ ⎩ t T ⎭ length ⎪⎭ ⎪⎩ ⎪⎩ ⎪⎭
(1)
Or, in exponent form, {k} = {m1 L1 t-3 T-1}. We obtain a value of k from a reference book. E.g. kcopper = 401 W/m⋅K. These units have dimensions of power/length⋅temperature. Since power is energy/time, we see immediately that Eq. 1 is correct. Alternatively, we can transform the units of k into primary units, Primary SI units of thermal conductivity:
kcopper = 401
W ⎛ N m ⎞⎛ kg m ⎞ kg ⋅ m = 401 3 2 ⎟ m K ⎜⎝ s W ⎟⎜ N s s ⋅K ⎠⎝ ⎠
(2)
Discussion We have used the principle of dimensional homogeneity to determine the primary dimensions of k. Namely, we utilized the fact that the dimensions of both terms of the given equation must be identical.
7-16 Solution
We are to determine the primary dimensions of each variable.
Analysis (a) Heat generation rate is energy per unit volume per unit time,
Primary dimensions of heat generation rate:
⎧ mL2 energy ⎫ ⎪⎪ t 2 { g } = ⎧⎨ ⎬=⎨ 3 ⎩ volume × time ⎭ ⎪ L t ⎩⎪
⎫ ⎪⎪ ⎧ m ⎫ ⎬=⎨ 3⎬ ⎪ ⎩ Lt ⎭ ⎭⎪
(1)
Or, in exponent form, { g } = {m1 L-1 t-3}. (b) Heat flux is energy per unit area per unit time,
Primary dimensions of heat flux:
⎧ mL2 energy ⎫ ⎪⎪ t 2 {q} = ⎧⎨ ⎬=⎨ 2 ⎩ area × time ⎭ ⎪ L t ⎩⎪
⎫ ⎪⎪ ⎧ m ⎫ ⎬=⎨ 3⎬ ⎪ ⎩t ⎭ ⎭⎪
(2)
Or, in exponent form, { q } = {m1 t-3}. (c) Heat flux is energy per unit area per unit time per unit temperature,
Primary dimensions of heat transfer coefficient:
⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ energy ⎪⎪ ⎧ m ⎫ t2 {h} = ⎨ ⎬=⎨ 2 ⎬=⎨ 3 ⎬ ⎩ area × time × temperature ⎭ ⎪ L × t × T ⎪ ⎩ t T ⎭ ⎪⎩ ⎪⎭
(3)
Or, in exponent form, {h} = {m1 t-3 T-1}. Discussion variables.
In the field of heat transfer it is critical that one be careful with the dimensions (and units) of heat transfer
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Chapter 7 Dimensional Analysis and Modeling 7-17 Solution dimensions.
We are to choose three properties or constants and write out their names, their SI units, and their primary
Analysis
There are many options. For example,
Students may choose cv (specific heat at constant volume). The units are kJ/kg⋅K, which is energy per mass per temperature. Thus, Primary dimensions of specific heat at constant volume: ⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ t 2 ⎪⎪ ⎧ L2 ⎫ energy {cv } = ⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩ mass × temperature ⎭ ⎪ m × T ⎪ ⎩ t T ⎭ ⎩⎪ ⎭⎪
(1)
Or, in exponent form, {cv} = { L2 t-2 T-1}. Students may choose v (specific volume). The units are m3/kg, which is volume per mass. Thus, volume ⎫ ⎧ L3 ⎫ ⎬=⎨ ⎬ ⎩ mass ⎭ ⎩ m ⎭
{v } = ⎧⎨
Primary dimensions of specific volume:
(2)
Or, in exponent form, {v} = {m-1 L3}. Students may choose hfg (latent heat of vaporization). The units are kJ/kg, which is energy per mass. Thus, Primary dimensions of latent heat of vaporization:
{h } fg
⎧ mL2 ⎧ energy ⎫ ⎪⎪ t 2 =⎨ ⎬=⎨ ⎩ mass ⎭ ⎪ m ⎪⎩
⎫ 2 ⎪⎪ ⎧ L ⎫ = ⎬ ⎨ 2⎬ ⎪ ⎩t ⎭ ⎪⎭
(3)
Or, in exponent form, {hfg} = {L2 t-2}. (The same dimensions hold for hf and hg.) Students may choose sf (specific entropy of a saturated liquid). The units are kJ/kg⋅K, which is energy per mass per temperature. Thus, Primary dimensions of specific entropy of a saturated liquid: ⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ t 2 ⎪⎪ ⎧ L2 ⎫ energy sf = ⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩ mass × temperature ⎭ ⎪ m × T ⎪ ⎩ t T ⎭ ⎪⎩ ⎪⎭
{ }
(4)
Or, in exponent form, {sf} = { L2 t-2 T-1}. (The same dimensions hold for sfg and sg.) Discussion
Students’ answers will vary. There are some other choices.
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Chapter 7 Dimensional Analysis and Modeling 7-18E Solution dimensions.
We are to choose three properties or constants and write out their names, their SI units, and their primary
Analysis
There are many options. For example,
Students may choose cv (specific heat at constant volume). The units are Btu/lbm⋅R, which is energy per mass per temperature. Thus, Primary dimensions of specific heat at constant volume: ⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ t 2 ⎪⎪ ⎧ L2 ⎫ energy {cv } = ⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩ mass × temperature ⎭ ⎪ m × T ⎪ ⎩ t T ⎭ ⎩⎪ ⎭⎪
(1)
Or, in exponent form, {cv} = { L2 t-2 T-1}. Students may choose v (specific volume). The units are ft3/lbm, which is volume per mass. Thus, volume ⎫ ⎧ L3 ⎫ ⎬=⎨ ⎬ ⎩ mass ⎭ ⎩ m ⎭
{v } = ⎧⎨
Primary dimensions of specific volume:
(2)
Or, in exponent form, {v} = {m-1 L3}. Students may choose hfg (latent heat of vaporization). The units are Btu/lbm, which is energy per mass. Thus, Primary dimensions of latent heat of vaporization:
{h } fg
⎧ mL2 ⎧ energy ⎫ ⎪⎪ t 2 =⎨ ⎬=⎨ ⎩ mass ⎭ ⎪ m ⎪⎩
⎫ 2 ⎪⎪ ⎧ L ⎫ = ⎬ ⎨ 2⎬ ⎪ ⎩t ⎭ ⎪⎭
(3)
Or, in exponent form, {hfg} = {L2 t-2}. (The same dimensions hold for hf and hg.) Students may choose sf (specific entropy of a saturated liquid). The units are Btu/lbm⋅R, which is energy per mass per temperature. Thus, Primary dimensions of specific entropy of a saturated liquid: ⎧ mL2 ⎫ ⎧ ⎫ ⎪⎪ t 2 ⎪⎪ ⎧ L2 ⎫ energy sf = ⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩ mass × temperature ⎭ ⎪ m × T ⎪ ⎩ t T ⎭ ⎪⎩ ⎪⎭
{ }
(4)
Or, in exponent form, {sf} = { L2 t-2 T-1}. (The same dimensions hold for sfg and sg.) Discussion
Students’ answers will vary. There are some other choices.
7-10 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling Dimensional Homogeneity
7-19C Solution
We are to explain the law of dimensional homogeneity.
Analysis The law of dimensional homogeneity states that every additive term in an equation must have the same dimensions. As a simple counter example, an equation with one term of dimensions length and another term of dimensions temperature would clearly violate the law of dimensional homogeneity – you cannot add length and temperature. All terms in the equation must have the same dimensions. Discussion If in the solution of an equation you realize that the dimensions of two terms are not equivalent, this is a sure sign that you have made a mistake somewhere!
7-20 Solution We are to determine the primary dimensions of the gradient operator, and then verify that primary dimensions of each additive term in the equation are the same. Analysis (a) By definition, the gradient operator is a three-dimensional derivative operator. For example, in Cartesian coordinates,
Gradient operator in Cartesian coordinates: G ⎛ ∂ ∂ ∂ ⎞ G ∂ G ∂ G ∂ ∇=⎜ , , ⎟=i +j +k ∂x ∂y ∂z ⎝ ∂x ∂y ∂z ⎠ Therefore its dimensions must be 1/length. Thus,
G
Primary dimensions of the gradient operator:
{∇} = ⎧⎨⎩ L1 ⎫⎬⎭
G Or, in exponent form, { ∇ } = { L-1}.
(b) Similarly, the primary dimensions of a time derivative (∂/∂t) are 1/time. Also, the primary dimensions of velocity are length/time, and the primary dimensions of acceleration are length/time2. Thus each term in the given equation can be written in terms of primary dimensions,
{a} = ⎧⎨
{a} = ⎨⎧ G
length ⎫ 2 ⎬ ⎩ time ⎭
G
⎧ length ⎫ G ⎪⎧ ∂V ⎪⎫ ⎪⎪ time ⎪⎪ ⎧ length ⎫ ⎨ ⎬=⎨ ⎬=⎨ 2 ⎬ ⎩⎪ ∂t ⎭⎪ ⎪ time ⎪ ⎩ time ⎭ ⎩⎪ ⎭⎪
L⎫ 2 ⎬ ⎩t ⎭
G ⎪⎧ ∂V ⎪⎫ ⎧ L ⎫ ⎨ ⎬=⎨ 2⎬ ⎩⎪ ∂t ⎭⎪ ⎩ t ⎭
1 length ⎫ ⎧ length ⎫ × × ⎬=⎨ ⎬ {(V ⋅∇ )V } = ⎨⎩⎧ length time length time ⎭ ⎩ time ⎭ G G G
2
{(V ⋅ ∇ )V } = ⎩⎧⎨ tL ⎭⎫⎬ G G G
2
Indeed, all three additive terms have the same dimensions, namely {L1 t-2}. Discussion If the dimensions of any of the terms were different from the others, it would be a sure sign that an error was made somewhere in deriving or copying the equation.
7-11 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling 7-21 Solution We are to determine the primary dimensions of each additive term in the equation, and we are to verify that the equation is dimensionally homogeneous. Analysis The primary dimensions of the time derivative (∂/∂t) are 1/time. The primary dimensions of the gradient vector are 1/length, and the primary dimensions of velocity are length/time. Thus each term in the equation can be written in terms of primary dimensions, ⎧ mass×length ⎫ G ⎪⎪ ⎧⎪ F ⎫⎪ ⎧ force ⎫ ⎪⎪ time 2 ⎨ ⎬=⎨ ⎬=⎨ ⎬ mass ⎪⎩ m ⎪⎭ ⎩ mass ⎭ ⎪ ⎪ ⎪⎩ ⎪⎭
G ⎪⎧ F ⎪⎫ ⎧ L ⎫ ⎨ ⎬=⎨ 2⎬ ⎩⎪ m ⎭⎪ ⎩ t ⎭
⎧ length ⎫ G ⎧⎪ ∂V ⎫⎪ ⎪⎪ time ⎪⎪ ⎨ ⎬=⎨ ⎬ ⎩⎪ ∂t ⎭⎪ ⎪ time ⎪ ⎪⎩ ⎪⎭
G ⎧⎪ ∂V ⎫⎪ ⎧ L ⎫ ⎨ ⎬=⎨ 2⎬ ⎩⎪ ∂t ⎭⎪ ⎩ t ⎭
1 length ⎫ × × ⎬ {(V ⋅∇ )V } = ⎧⎨⎩ length time length time ⎭
{(V ⋅∇ )V } = ⎧⎨⎩ tL ⎫⎬⎭
G G G
G G G
2
Indeed, all three additive terms have the same dimensions, namely {L1 t-2}. Discussion
The dimensions are, in fact, those of acceleration.
7-22 Solution We are to determine the primary dimensions of each additive term in Eq. 1, and we are to verify that the equation is dimensionally homogeneous. Analysis The primary dimensions of the material derivative (D/Dt) are 1/time. The primary dimensions of volume are length3, and the primary dimensions of velocity are length/time. Thus each term in the equation can be written in terms of primary dimensions, length 3 ⎫ ⎧ 1 ⎫ ⎧ 1 DV ⎫ ⎧ 1 = × ⎨ ⎬ ⎨ ⎬=⎨ ⎬ 3 time ⎭ ⎩ time ⎭ ⎩V Dt ⎭ ⎩ length
⎧ 1 DV ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩V Dt ⎭ ⎩ t ⎭
⎧ length ⎫ ⎧ ∂u ⎫ ⎪⎪ time ⎪⎪ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬=⎨ ⎬ ⎩ ∂x ⎭ ⎪ length ⎪ ⎩ time ⎭ ⎩⎪ ⎭⎪
⎧ ∂u ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩ ∂x ⎭ ⎩ t ⎭
⎧ length ⎫ ⎧ ∂v ⎫ ⎪⎪ time ⎪⎪ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬=⎨ ⎬ ⎩ ∂y ⎭ ⎪ length ⎪ ⎩ time ⎭ ⎩⎪ ⎭⎪
⎧ ∂v ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩ ∂y ⎭ ⎩ t ⎭
⎧ length ⎫ ⎧ ∂w ⎫ ⎪⎪ time ⎪⎪ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬=⎨ ⎬ ⎩ ∂z ⎭ ⎪ length ⎪ ⎩ time ⎭ ⎪⎩ ⎪⎭
⎧ ∂w ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩ ∂z ⎭ ⎩ t ⎭
Indeed, all four additive terms have the same dimensions, namely {t-1}. Discussion If the dimensions of any of the terms were different from the others, it would be a sure sign that an error was made somewhere in deriving or copying the equation.
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Chapter 7 Dimensional Analysis and Modeling 7-23 Solution We are to determine the primary dimensions of each additive term, and we are to verify that the equation is dimensionally homogeneous. Analysis The primary dimensions of the velocity components are length/time. The primary dimensions of coordinates r and z are length, and the primary dimensions of coordinate θ are unity (it is a dimensionless angle). Thus each term in the equation can be written in terms of primary dimensions, length ⎫ ⎧ length ⎧⎪ 1 ∂ ( rur ) ⎫⎪ ⎪⎪ 1 time ⎪⎪ = ⎧ 1 ⎫ × ⎨ ⎬=⎨ ⎬ ⎨ ⎬ length ⎪ ⎩ time ⎭ ⎩⎪ r ∂r ⎭⎪ ⎪ length ⎩⎪ ⎭⎪
⎧⎪ 1 ∂ ( rur ) ⎫⎪ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩⎪ r ∂r ⎭⎪ ⎩ t ⎭
length ⎫ ⎧ ⎧ 1 ∂uθ ⎫ ⎪⎪ 1 ⎪⎪ ⎧ 1 ⎫ × time ⎬ = ⎨ ⎨ ⎬ == ⎨ ⎬ 1 ⎪ ⎩ time ⎭ ⎩ r ∂θ ⎭ ⎪ length ⎪⎩ ⎪⎭
⎧ 1 ∂uθ ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩ r ∂θ ⎭ ⎩ t ⎭
⎧ length ⎫ ⎧ ∂u z ⎫ ⎪⎪ time ⎪⎪ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬=⎨ ⎬ ⎩ ∂z ⎭ ⎪ length ⎪ ⎩ time ⎭ ⎩⎪ ⎭⎪
⎧ ∂u z ⎫ ⎧ 1 ⎫ ⎨ ⎬=⎨ ⎬ ⎩ ∂z ⎭ ⎩ t ⎭
Indeed, all three additive terms have the same dimensions, namely {t-1}. Discussion If the dimensions of any of the terms were different from the others, it would be a sure sign that an error was made somewhere in deriving or copying the equation.
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Chapter 7 Dimensional Analysis and Modeling 7-24 Solution We are to determine the primary dimensions of each additive term in the equation, and we are to verify that the equation is dimensionally homogeneous. Analysis The primary dimensions of heat transfer rate are energy/time. The primary dimensions of mass flow rate are mass/time, and those of specific heat are energy/mass⋅temperature, as found in Problem 7-14. Thus each term in the equation can be written in terms of primary dimensions, ⎧ mL2 ⎧ energy ⎫ ⎪⎪ t 2 Q = ⎨ ⎬=⎨ ⎩ time ⎭ ⎪ t ⎩⎪
{}
⎫ 2 ⎪⎪ ⎧ mL ⎫ ⎬=⎨ 3 ⎬ ⎪ ⎩ t ⎭ ⎭⎪
⎧
⎫
{Q } = ⎨ mL ⎬ t ⎩
2
3
⎭
T } {mC
⎧ ⎫ mL2 2 2 ⎧ mass ⎫ ⎪⎪ m energy ⎪⎪ ⎧ mL ⎫ =⎨ × × temperature ⎬ = ⎨ × t × T ⎬ = ⎨ 3 ⎬ ⎩ time mass × temperature ⎭ ⎪ t m×T ⎪ ⎩ t ⎭ ⎪⎩ ⎪⎭
mL T }=⎨ {mC ⎬ t
T } {mC
⎧ ⎫ mL2 ⎪ ⎧ mL2 ⎫ 2 ⎧ mass ⎫ ⎪⎪ m energy ⎪ =⎨ × × temperature ⎬ = ⎨ × t × T ⎬ = ⎨ 3 ⎬ ⎩ time mass × temperature ⎭ ⎪ t m×T ⎪ ⎩ t ⎭ ⎩⎪ ⎭⎪
mL T }=⎨ {mC ⎬ t
p out
p in
⎧
p out
⎩
2
3
⎧
p in
⎩
⎫ ⎭
2
3
⎫ ⎭
Indeed, all three additive terms have the same dimensions, namely {m1 L2 t-3}. Discussion We could also have left the temperature difference in parentheses as a temperature difference (same dimensions as the individual temperatures), and treated the equation as having only two terms.
7-25 Solution We are to determine the primary dimensions of each additive term, and we are to verify that the equation is dimensionally homogeneous. Analysis The primary dimensions of the time derivative (d/dt) are 1/time. The primary dimensions of density are mass/length3, those of volume are length3, those of area are length2, and those of velocity are length/time. The primary G G dimensions of unit vector n are unity, i.e. {1} (in other words n has no dimensions). Finally, the primary dimensions of b, which is defined as B per unit mass, are {B/m}. Thus each term in the equation can be written in terms of primary dimensions, ⎧ dBsys ⎫ ⎧ B ⎫ ⎨ ⎬=⎨ ⎬ ⎩ dt ⎭ ⎩ time ⎭ ⎧d ⎨ ⎩ dt
{∫
CS
∫
CV
⎧ dBsys ⎫ ⎧ B ⎫ ⎨ ⎬=⎨ ⎬ ⎩ dt ⎭ ⎩ t ⎭
⎧ 1 ⎫ mass B × × × length 3 ⎬ 3 ⎩ time length mass ⎭
⎫
ρ bdV ⎬ = ⎨
G G
⎭
}
⎧ mass ⎫ B length × × × 1× length 2 ⎬ 3 ⎩ length mass time ⎭
ρ bVr ⋅ ndA = ⎨
⎧d ⎨ ⎩ dt
∫
{∫
ρ bVr ⋅ ndA = ⎨ ⎬
CS
CV
⎫ ⎭
⎧B⎫ ⎩t⎭
}
⎧B⎫ ⎩t⎭
ρ bdV ⎬ = ⎨ ⎬ G G
Indeed, all three additive terms have the same dimensions, namely {B t-1}. Discussion
The RTT for property B has dimensions of rate of change of B.
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Chapter 7 Dimensional Analysis and Modeling 7-26 Solution We are to determine the primary dimensions of the first three additive term, and we are to verify that those terms are dimensionally homogeneous. Then we are to evaluate the dimensions of the adsorption coefficient. Analysis The primary dimensions of the time derivative (d/dt) are 1/time. Those of As are length2, those of V are length3, those of c are mass/length3, and those of V are length3/time. Thus the primary dimensions of the first three terms are ⎧ ⎨V ⎩
mass ⎫ ⎧ 3 ⎪ ⎪ dc ⎫ ⎪ ⎧ mass ⎫ 3 length ⎪ ⎬ = ⎨length ⎬=⎨ ⎬ time ⎪ ⎩ time ⎭ dt ⎭ ⎪ ⎪⎭ ⎩⎪
{S} = ⎧⎨
{V c} = ⎧⎨⎩ length time
3
⎧ dc ⎫ ⎧ m ⎫ ⎨V ⎬ = ⎨ ⎬ ⎩ dt ⎭ ⎩ t ⎭
mass ⎫ ⎬ ⎩ time ⎭
{S} = ⎧⎨
mass ⎫ ⎧ mass ⎫ ⎬=⎨ ⎬ length 3 ⎭ ⎩ time ⎭
{V c} = ⎧⎨⎩ mt ⎫⎬⎭
×
m⎫ ⎬ ⎩t ⎭
Indeed, the first three additive terms have the same dimensions, namely {m1 t-1}. Since the equation must be dimensionally homogeneous, the last term must have the same dimensions as well. We use this fact to find the dimensions of kw, mass ⎫ {cAs kw } = ⎧⎨ ⎬ ⎩ time ⎭
⎫ mass ⎧ mass ⎫ ⎧ ⎪⎪ time ⎪⎪ ⎪⎪ ⎪⎪ time {k w } = ⎨ ⎬=⎨ ⎬ ⎪ cAs ⎪ ⎪ mass 3 × length 2 ⎪ ⎩⎪ ⎭⎪ ⎩⎪ length ⎭⎪
{kw } = ⎧⎨
L⎫ ⎬ ⎩t⎭
Or, in exponent form, {kw} = {L1 t-1}. The dimensions of wall adsorption coefficient are those of velocity. Discussion
In fact, some authors call kw a “deposition velocity”.
7-15 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling Nondimensionalization of Equations
7-27C Solution Analysis problem.
We are to give the primary reason for nondimensionalizing an equation. The primary reason for nondimensionalizing an equation is to reduce the number of parameters in the
Discussion As shown in the examples in the text, nondimensionalization of an equation reduces the number of independent parameters in the problem, simplifying the analysis.
7-28 Solution
We are to nondimensionalize all the variables, and then re-write the equation in nondimensionalized form.
Assumptions
1 The air in the room is well mixed so that c is only a function of time.
Analysis (a) We nondimensionalize the variables by inspection according to their dimensions,
Nondimensionalized variables: A V c V L2 S , t* = t 3 , As * = 2s , k w * = k w , and S * = V * = 3 , c* = climit L L L V climitV (b) We substitute these into the equation to generate the nondimensionalized equation, V *L3
d ( c*climit )
− V c*c − ( c*c ) A *L2 ⎛⎜ k * V = S * c V s w limit limit limit ⎜ L2 ⎛ L3 ⎞ ⎝ d ⎜ t* ⎟ ⎝ V ⎠
(
)
⎞ ⎟⎟ ⎠
(1)
We notice that every term in Eq. 1 contains the quantity V climit . We divide every term by this quantity to get a nondimensionalized form of the equation, Nondimensionalized equation:
V*
dc* = S * − c* − c*As *kw * dt*
No dimensionless groups have arisen in this nondimensionalization. Discussion Since all the characteristic scales disappear, no dimensionless groups have arisen. Since there are no dimensionless parameters, one solution in nondimensionalized variables is valid for all combinations of L, V , and climit.
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Chapter 7 Dimensional Analysis and Modeling 7-29 Solution We are to nondimensionalize the equation, and identify the dimensionless parameters that appear in the nondimensionalized equation. Assumptions
1 The flow is steady. 2 The flow is incompressible.
Analysis We plug the nondimensionalized variables into the equation. For example, u = u*U and x = x*L in the first term. The result is U ∂u * U ∂v * U ∂w * + + =0 L ∂x * L ∂y * L ∂z *
or, after simplifying, Nondimensionalized incompressible flow relationship: ∂u * ∂v * ∂w * + + =0 ∂x * ∂y * ∂z *
(1)
There are no nondimensional parameters in the nondimensionalized equation. The original equation comes from pure kinematics – there are no fluid properties involved in the equation, and therefore it is not surprising that no nondimensional parameters appear in the nondimensionalized form of the equation, Eq. 1. Discussion We show in Chap. 9 that the equation given in this problem is the differential equation for conservation of mass for an incompressible flow field – the incompressible continuity equation.
7-30 Solution We are to nondimensionalize the equation of motion and identify the dimensionless parameters that appear in the nondimensionalized equation. Analysis is
We plug in the nondimensionalized variables. For example, u = u*V and x = x*L in the first term. The result L3 f 1 DV * V ∂u * V ∂v * V ∂w * = + + L3 V * Dt * L ∂x * L ∂y * L ∂z *
or, after simplifying, ⎛ fL ⎞ 1 DV * ∂u * ∂v * ∂w * = + + ⎜ ⎟ ⎝ V ⎠ V * Dt * ∂x * ∂y * ∂z *
(1)
We recognize the nondimensional parameter in parentheses in Eq. 1 as St, the Strouhal number. We can re-write Eq. 1 as Nondimensionalized oscillating compressible flow relationship: 1 DV * ∂u * ∂v * ∂w * St = + + V * Dt * ∂x * ∂y * ∂z * Discussion We show in Chap. 9 that the given equation of motion is the differential equation for conservation of mass for an unsteady, compressible flow field – the general continuity equation. We may also use angular frequency ω (radians per second) in place of physical frequency f (cycles per second), with the same result.
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Chapter 7 Dimensional Analysis and Modeling 7-31 Solution We are to determine the primary dimensions of the stream function, nondimensionalize the variables, and then re-write the definition of ψ in nondimensionalized form. Assumptions
1 The flow is incompressible. 2 The flow is two-dimensional in the x-y plane.
Analysis ψ,
(a) We use that fact that all equations must be dimensionally homogeneous. We solve for the dimensions of
Primary dimensions of stream function:
2 L ⎫ ⎧L ⎫ × L⎬ = ⎨ ⎬ ⎩t ⎭ ⎩ t ⎭
{ψ } = {u} × { y} = ⎨⎧
Or, in exponent form, {ψ} = {L2 t-1}. (b) We nondimensionalize the variables by inspection according to their dimensions, Nondimensionalized variables: x y x* = y* = L L
u* = u
t L
v* = v
t L
ψ* =ψ
t L2
(c) We generate the nondimensionalized equations, ⎛ L2 ⎞ ∂ψ * ⎜ ⎟ ⎛ L⎞ ⎝ t ⎠ u *⎜ ⎟ = t y * ∂ ( L) ⎝ ⎠
⎛ L2 ⎞ ∂ψ * ⎜ ⎟ ⎛ L⎞ ⎝ t ⎠ v *⎜ ⎟ = − t x * ∂ ( L) ⎝ ⎠
We notice that every term in both parts of the above equation contains the ratio L/t. We divide every term by L/t to get the final nondimensionalized form of the equations, Nondimensionalized stream function equations:
u* =
∂ψ * ∂y *
v* = −
∂ψ * ∂x *
No dimensionless groups have arisen in this nondimensionalization. Discussion
Since all the nondimensionalized variables scale with L and t, no dimensionless groups have arisen.
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Chapter 7 Dimensional Analysis and Modeling 7-32 Solution We are to nondimensionalize the equation of motion and identify the dimensionless parameters that appear in the nondimensionalized equation. G G Analysis We plug the nondimensionalized variables into the equation. For example, t = t*/ω and V = V∞V * in the first term on the right hand side. The result is G G ∂V * U 2 G G G 2 V * ⋅∇ * V * + ω L F / m * = ωU L ∂t *
(
)
(
)
or, after simplifying by multiplying each term by L/ V∞2, G 2 G G G ⎛ωL ⎞ G ⎛ ω L ⎞ ∂V * + V * ⋅∇ * V * ⎜ ⎟ F /m *= ⎜ ⎟ ⎝ V∞ ⎠ ⎝ V∞ ⎠ ∂t *
(
)
(
)
(1)
We recognize the nondimensional parameter in parentheses in Eq. 1 as St, the Strouhal number. We re-write Eq. 1 as Nondimensionalized Newton’s second law for incompressible oscillatory G G G G G ∂V * 2 + V * ⋅∇ * V * flow: (St ) F / m * = (St ) ∂t *
(
)
(
)
Discussion We used angular frequency ω in this problem. The same result would be obtained if we used physical frequency. Equation 1 is the basis for forming the differential equation for conservation of linear momentum for an unsteady, incompressible flow field.
7-33 Solution
We are to nondimensionalize the Bernoulli equation and generate an expression for the pressure coefficient.
Assumptions other terms.
1 The flow is incompressible. 2 Gravitational terms in the Bernoulli equation are negligible compared to the
Analysis
We nondimensionalize the equation by dividing each term by the dynamic pressure,
Nondimensionalization:
P 1 ρV∞ 2 2
+
1 ρV∞ 2 , 2
P∞ V2 = +1 2 1 V∞ ρV∞ 2 2
Rearranging, Pressure coefficient:
Cp =
P − P∞ V2 = 1− 2 1 V∞ ρV∞ 2 2
Discussion Pressure coefficient is a useful dimensionless parameter that is inversely related to local air speed – as local air speed V increases, Cp decreases.
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Chapter 7 Dimensional Analysis and Modeling
Dimensional Analysis and Similarity
7-34C Solution
We are to list the three primary purposes of dimensional analysis.
Analysis The three primary purposes of dimensional analysis are: 1. To generate nondimensional parameters that help in the design of experiments and in the reporting of experimental results. 2. To obtain scaling laws so that prototype performance can be predicted from model performance. 3. To (sometimes) predict trends in the relationship between parameters. Discussion
Dimensional analysis is most useful for difficult problems that cannot be solved analytically.
7-35C Solution prototype.
We are to list and describe the three necessary conditions for complete similarity between a model and a
Analysis The three necessary conditions for complete similarity between a model and a prototype are: 1. Geometric similarity – the model must be the same shape as the prototype, but scaled by some constant scale factor. 2. Kinematic similarity – the velocity at any point in the model flow must be proportional (by a constant scale factor) to the velocity at the corresponding point in the prototype flow. 3. Dynamic similarity – all forces in the model flow scale by a constant factor to corresponding forces in the prototype flow. Discussion
Complete similarity is achievable only when all three of the above similarity conditions are met.
7-36 Solution For a scale model of a submarine being tested in air, we are to calculate the wind tunnel speed required to achieve similarity with the prototype submarine that moves through water at a given speed. Assumptions 1 Compressibility of the air is assumed to be negligible. 2 The wind tunnel walls are far enough away so as to not interfere with the aerodynamic drag on the model sub. 3 The model is geometrically similar to the prototype. Properties For water at T = 15oC and atmospheric pressure, ρ = 999.1 kg/m3 and μ = 1.138 × 10-3 kg/m⋅s. For air at T = o 25 C and atmospheric pressure, ρ = 1.184 kg/m3 and μ = 1.849 × 10-5 kg/m⋅s. Analysis
Similarity is achieved when the Reynolds number of the model is equal to that of the prototype, Re m =
Similarity:
ρ pVp Lp ρ mVm Lm = Re p = μm μp
(1)
We solve Eq. 1 for the unknown wind tunnel speed, ⎛μ Vm = Vp ⎜ m ⎜ μp ⎝
⎞ ⎛ ρ p ⎞ ⎛ Lp ⎞ ⎟⎟ ⎜ ⎟⎜ ⎟ ⎠ ⎝ ρ m ⎠ ⎝ Lm ⎠ ⎛ 1.849 × 10−5 kg/m ⋅ s ⎞ ⎛ 999.1 kg/m3 ⎞ 8 = 61.4 m/s = ( 0.560 m/s ) ⎜ ⎟⎜ 3 ⎟( ) −3 ⎝ 1.138 × 10 kg/m ⋅ s ⎠ ⎝ 1.184 kg/m ⎠
Discussion At this air temperature, the speed of sound is around 346 m/s. Thus the Mach number in the wind tunnel is equal to 61.4/346 = 0.177. This is sufficiently low that the incompressible flow approximation is reasonable.
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Chapter 7 Dimensional Analysis and Modeling 7-37 Solution For a scale model of a submarine being tested in air, we are to calculate the wind tunnel speed required to achieve similarity with the prototype submarine that moves through water at a given speed. Assumptions 1 Compressibility of the air is assumed to be negligible. 2 The wind tunnel walls are far enough away so as to not interfere with the aerodynamic drag on the model sub. 3 The model is geometrically similar to the prototype. Properties For water at T = 15oC and atmospheric pressure, ρ = 999.1 kg/m3 and μ = 1.138 × 10-3 kg/m⋅s. For air at T = o 25 C and atmospheric pressure, ρ = 1.184 kg/m3 and μ = 1.849 × 10-5 kg/m⋅s. Analysis
Similarity is achieved when the Reynolds number of the model is equal to that of the prototype, Rem =
Similarity:
ρ pVp Lp ρ mVm Lm = Re p = μm μp
(1)
We solve Eq. 1 for the unknown wind tunnel speed, ⎛μ Vm = Vp ⎜ m ⎜ ⎝ μp
⎞ ⎛ ρ p ⎞ ⎛ Lp ⎞ ⎛ 1.849 × 10 −5 kg/m ⋅ s ⎞⎛ 999.1 kg/m 3 ⎞ = 0.560 m/s ⎜ 24 ) = 184 m/s ⎟⎜ ⎟ ⎜ ⎟ ⎟⎜ 3 ⎟( −3 ⎟ ⎝ 1.138 × 10 kg/m ⋅ s ⎠⎝ 1.184 kg/m ⎠ ⎠ ⎝ ρ m ⎠ ⎝ Lm ⎠
At this air temperature, the speed of sound is around 346 m/s. Thus the Mach number in the wind tunnel is equal to 184/346 = 0.532. The Mach number is sufficiently high that the incompressible flow approximation is not reasonable. The wind tunnel should be run at a flow speed at which the Mach number is less than one-third of the speed of sound. At this lower speed, the Reynolds number of the model will be too small, but the results may still be usable, either by extrapolation to higher Re, or if we are fortunate enough to have Reynolds number independence, as discussed in Section 7-5. Discussion
It is also unlikely that a small instructional wind tunnel can achieve such a high speed.
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Chapter 7 Dimensional Analysis and Modeling 7-38E Solution For a prototype parachute and its model we are to calculate drag coefficient, and determine the wind tunnel speed that ensures dynamic similarity. Then we are to estimate the aerodynamic drag on the model. Assumptions
1 The model is geometrically similar to the prototype.
Properties
For air at 60oF and standard atmospheric pressure, ρ = 0.07633 lbm/ft3 and μ = 1.213 × 10-5 lbm/ft⋅s.
Analysis (a) The aerodynamic drag on the prototype parachute is equal to the total weight. We can then easily calculate the drag coefficient CD,
Drag coefficient:
CD =
FD ,p 1 2
ρ pVp 2 Ap
230 lbf
= 1 2
( 0.07633 lbm/ft ) ( 20 ft/s ) π ( 3
2
24 ft )
2
⎛ 32.2 lbm ft ⎞ ⎜ lbf s 2 ⎟ = 1.07 ⎝ ⎠
4
(b) We must match model and prototype Reynolds numbers in order to achieve dynamic similarity, Rem =
Similarity:
ρ pVp Lp ρ mVm Lm = Re p = μm μp
(1)
We solve Eq. 1 for the unknown wind tunnel speed, Wind tunnel speed:
⎛μ Vm = Vp ⎜ m ⎜ μp ⎝
⎞ ⎛ ρ p ⎞ ⎛ Lp ⎞ ⎟⎟ ⎜ ⎟⎜ ⎟ = ( 20 ft/s )(1)(1)(12 ) = 240 ft/s ⎠ ⎝ ρ m ⎠ ⎝ Lm ⎠
(2)
(c) As discussed in the text, if the fluid is the same and dynamic similarity between the model and the prototype is achieved, the aerodynamic drag force on the model is the same as that on the prototype. Thus, Aerodynamic drag on model:
FD ,m = FD ,p = 230 lbf
(3)
Discussion We should check that the wind tunnel speed of Eq. 2 is not too high that the incompressibility approximation becomes invalid. The Mach number at this speed is about 240/1120 = 0.214. Since this is less than 0.3, compressibility is not an issue in this model test. The drag force on the model is quite large, and a fairly hefty drag balance must be available to measure such a large force.
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Chapter 7 Dimensional Analysis and Modeling 7-39 Solution
We are to discuss why one would pressurize a wind tunnel.
Analysis As we see in some of the example problems and homework problems in this chapter, it is often difficult to achieve a high-enough wind tunnel speed to match the Reynolds number between a small model and a large prototype. Even if we were able to match the speed, the Mach number would often be too high. A pressurized wind tunnel has higher density air. At the same Reynolds number, the larger density leads to a lower air speed requirement. In other words, a pressurized wind tunnel can achieve higher Reynolds numbers for the same scale model. If the pressure were to be increased by a factor of 1.5, the air density would also go up by a factor of 1.5 (ideal gas law), assuming that the air temperature remains constant. Then the Reynolds number, Re = ρVL/μ, would go up by approximately 1.5. Note that we are also assuming that the viscosity does not change significantly with pressure, which is a reasonable assumption. Discussion The speed of sound is not a strong function of pressure, so Mach number is not affected significantly by pressurizing the wind tunnel. However, the power requirement for the wind tunnel blower increases significantly as air density is increased, so this must be taken into account when designing the wind tunnel.
7-40 Solution We are to estimate the drag on a prototype submarine in water, based on aerodynamic drag measurements performed in a wind tunnel. Assumptions 1 The model is geometrically similar. 2 The wind tunnel is run at conditions which ensure similarity between model and prototype. Properties For water at T = 15oC and atmospheric pressure, ρ = 999.1 kg/m3 and μ = 1.138 × 10-3 kg/m⋅s. For air at T = o 25 C and atmospheric pressure, ρ = 1.184 kg/m3 and μ = 1.849 × 10-5 kg/m⋅s. Analysis Since the Reynolds numbers have been matched, the nondimensionalized drag coefficient of the model equals that of the prototype, FD ,m
ρ mVm Lm 2
2
=
FD ,p
ρ pVp 2 Lp 2
(1)
We solve Eq. 1 for the unknown aerodynamic drag force on the prototype, FD,p, 2 ⎛ ρ p ⎞ ⎛ Vp ⎞ ⎛ Lp ⎞ ⎛ 999.1 kg/m ⋅ s ⎞ ⎛ 0.560 m/s ⎞ 2 = FD ,m ⎜ (8) = 10.3 N ⎟⎜ ⎟ ⎜ ⎟ = ( 2.3 N ) ⎜ ⎟⎜ ⎟ ⎝ 1.184 kg/m ⋅ s ⎠ ⎝ 61.4 m/s ⎠ ⎝ ρ m ⎠ ⎝ Vm ⎠ ⎝ Lm ⎠ 2
FD ,p
2
where we have used the wind tunnel speed calculated in Problem 7-36. Discussion Although the prototype moves at a much slower speed than the model, the density of water is much higher than that of air, and the prototype is eight times larger than the model. When all of these factors are combined, the drag force on the prototype is much larger than that on the model.
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Chapter 7 Dimensional Analysis and Modeling 7-41E Solution
The concept of similarity will be utilized to determine the speed of the wind tunnel.
Assumptions 1 Compressibility of the air is ignored (the validity of this assumption will be discussed later). 2 The wind tunnel walls are far enough away so as to not interfere with the aerodynamic drag on the model car. 3 The model is geometrically similar to the prototype. 4 Both the air in the wind tunnel and the air flowing over the prototype car are at standard atmospheric pressure. Properties
For air at T = 25oC and atmospheric pressure, ρ = 1.184 kg/m3 and μ = 1.849 × 10-5 kg/m⋅s.
Analysis Since there is only one independent Π in this problem, similarity is achieved if Π2,m = Π2,p, where Π2 is the Reynolds number. Thus, we can write Π 2,m = Re m =
ρ pVp Lp ρ mVm Lm = Π 2,p = Re p = μm μp
which can be solved for the unknown wind tunnel speed for the model tests, Vm, ⎛μ Vm = Vp ⎜ m ⎜ μp ⎝
⎞ ⎛ ρ p ⎞ ⎛ Lp ⎞ ⎟⎟ ⎜ ⎟⎜ ⎟ = ( 60.0 mph )(1)(1)( 4 ) = 240 mph ⎠ ⎝ ρ m ⎠ ⎝ Lm ⎠
Thus, to ensure similarity, the wind tunnel should be run at 240 miles per hour (to three significant digits). Discussion This speed is quite high, and the wind tunnel may not be able to run at that speed. We were never given the actual length of either car, but the ratio of Lp to Lm is known because the prototype is four times larger than the scale model. The problem statement contains a mixture of SI and English units, but it does not matter since we use ratios in the algebra.
7-42E Solution wind tunnel.
We are to estimate the drag on a prototype car, based on aerodynamic drag measurements performed in a
Assumptions 1 The model is geometrically similar. 2 The wind tunnel is run at conditions which ensure similarity between model and prototype. Properties
For air at T = 25oC and atmospheric pressure, ρ = 1.184 kg/m3 and μ = 1.849 × 10-5 kg/m⋅s.
Analysis Following the example in the text, since the Reynolds numbers have been matched, the nondimensionalized drag coefficient of the model equals that of the prototype, FD ,m
ρ mVm Lm 2
2
=
FD ,p
ρ pVp 2 Lp 2
(1)
We solve Eq. 1 for the unknown aerodynamic drag force on the prototype, FD,p, ⎛ ρ p ⎞ ⎛ Vp ⎞ ⎛ Lp ⎞ ⎛ 60.0 mph ⎞ 2 = FD ,m ⎜ ⎟⎜ ⎟ ⎜ ⎟ = ( 36.5 lbf )(1) ⎜ ⎟ ( 4 ) = 36.5 lbf ⎝ 240 mph ⎠ ⎝ ρ m ⎠ ⎝ Vm ⎠ ⎝ Lm ⎠ 2
FD ,p
2
2
where we have used the wind tunnel speed calculated in the previous problem. Discussion Since the air properties of the wind tunnel are identical to those of the air flowing around the prototype car, it turns out that the aerodynamic drag force on the prototype is the same as that on the model. This would not be the case if the wind tunnel air were at a different temperature or pressure compared to that of the prototype.
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Chapter 7 Dimensional Analysis and Modeling 7-43 Solution We are to discuss whether cold or hot air in a wind tunnel is better, and we are to support our answer by comparing air at two given temperatures. Properties For air at atmospheric pressure and at T = 10oC, ρ = 1.246 kg/m3 and μ = 1.778 × 10-5 kg/m⋅s. At T = 50oC, 3 ρ = 1.092 kg/m and μ = 1.963 × 10-5 kg/m⋅s. Analysis As we see in some of the example problems and homework problems in this chapter, it is often difficult to achieve a high-enough wind tunnel speed to match the Reynolds number between a small model and a large prototype. Even if we were able to match the speed, the Mach number would often be too high. Cold air has higher density than warm air. In addition, the viscosity of cold air is lower than that of hot air. Thus, at the same Reynolds number, the colder air leads to a lower air speed requirement. In other words, a cold wind tunnel can achieve higher Reynolds numbers than can a hot wind tunnel for the same scale model, all else being equal. We support our conclusion by comparing air at two temperatures,
Comparison of Reynolds numbers:
ρcoldVL Re cold μcold ρ μ 1.246 kg/m3 1.963 × 10−5 kg/m ⋅ s = = cold hot = = 1.26 ρ hotVL Re hot ρ hot μcold 1.092 kg/m3 1.778 × 10−5 kg/m ⋅ s μhot
Thus we see that the colder wind tunnel can achieve approximately 26% higher Reynolds number, all else being equal. Discussion There are other issues however. First of all, the denser air of the cold wind tunnel is harder to pump – the cold wind tunnel may not be able to achieve the same wind speed as the hot wind tunnel. Furthermore, the speed of sound is proportional to the square root of temperature. Thus, at colder temperatures, the Mach number is higher than at warmer temperatures for the same value of V, and compressibility effects are therefore more significant at lower temperatures.
7-44 Solution We are to calculate the speed and angular velocity (rpm) of a spinning baseball in a water channel such that flow conditions are dynamically similar to that of the actual baseball moving and spinning in air. Properties For air at T = 20oC and atmospheric pressure, ρ = 1.204 kg/m3 and μ = 1.825 × 10-5 kg/m⋅s. For water at T = o 20 C and atmospheric pressure, ρ = 998.0 kg/m3 and μ = 1.002 × 10-3 kg/m⋅s. Analysis The model (in the water) and the prototype (in the air) are actually the same baseball, so their characteristic lengths are equal, Lm = Lp. We match Reynolds number, Re m =
ρ pVp Lp ρ mVm Lm = Re p = μm μp
(1)
and solve for the required water tunnel speed for the model tests, Vm, ⎛μ Vm = Vp ⎜ m ⎜ μp ⎝
⎞ ⎛ ρ p ⎞⎛ Lp ⎞ ⎛ 1.002 × 10 −3 kg/m ⋅ s ⎞ ⎛ 1.204 kg/m3 ⎞ 1 = 5.30 mph = ( 80.0 mph ) ⎜ ⎟⎜ ⎟⎜ ⎟ ⎟⎜ 3 ⎟( ) −5 ⎟ ρ m Lm ⎝ 1.825 × 10 kg/m ⋅ s ⎠ ⎝ 998.0 kg/m ⎠ ⎠⎝ ⎠ ⎠⎝
(2)
We also match Strouhal numbers, recognizing that n is proportional to f, St m =
f p Lp f m Lm = St p = Vm Vp
→
nm Lm np Lp = Vm Vp
(3)
from which we solve for the required spin rate in the water tunnel, ⎛ Lp nm = np ⎜ ⎝ Lm
⎞ ⎛ Vm ⎟ ⎜⎜ ⎠ ⎝ Vp
⎞ ⎛ 5.30 mph ⎞ ⎟ = ( 300 rpm )(1) ⎜ ⎟ = 19.9 rpm ⎟ ⎝ 80.0 mph ⎠ ⎠
(4)
Discussion Because of the difference in fluid properties between air and water, the required water tunnel speed is much lower than that in air. In addition, the spin rate is much lower, making flow visualization easier.
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Chapter 7 Dimensional Analysis and Modeling
Dimensionless Parameters and the Method of Repeating Variables
7-45 Solution
We are to verify that the Archimedes number is dimensionless.
Analysis
Archimedes number is defined as
Archimedes number:
Ar =
ρ s gL3 ( ρs − ρ ) μ2
(1)
We know the primary dimensions of density, gravitational acceleration, length, and viscosity. Thus,
Primary dimensions of Archimedes number:
⎧m L 3 ⎫ ⎪ 3 2 L m⎪ {Ar} = ⎪⎨ L t 2 3 ⎪⎬ = {1} L ⎪ ⎪ m ⎪⎭ ⎩⎪ L2 t 2
(2)
Discussion If the primary dimensions were not unity, we would assume that we made an error in the dimensions of one or more of the parameters.
7-46 Solution
We are to verify that the Grashof number is dimensionless.
Analysis
Grashof number is defined as
Grashof number:
Gr =
g β ΔT L3 ρ 2
μ2
(1)
We know the primary dimensions of density, gravitational acceleration, length, temperature, and viscosity. The dimensions of coefficient of thermal expansion β are 1/temperature. Thus,
Primary dimensions of Grashof number:
⎧ L 1 3 m2 ⎫ ⎪⎪ 2 TL 6 ⎪⎪ Gr = { } ⎨ t T 2 L ⎬ = {1} m ⎪ ⎪ ⎪⎭ L2 t 2 ⎩⎪
(2)
Discussion If the primary dimensions were not unity, we would assume that we made an error in the dimensions of one or more of the parameters.
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Chapter 7 Dimensional Analysis and Modeling 7-47 Solution We are to verify that the Rayleigh number is dimensionless, and determine what other established nondimensional parameter is formed by the ratio of Ra and Gr. Analysis
Rayleigh number is defined as Ra =
Rayleigh number:
g β ΔT L3 ρ 2 c p kμ
(1)
We know the primary dimensions of density, gravitational acceleration, length, temperature, and viscosity. The dimensions of coefficient of thermal expansion β are 1/temperature, those of specific heat cp are length2/time2⋅temperature (Problem 714), and those of thermal conductivity k are mass⋅length/time3⋅temperature. Thus,
Primary dimensions of Rayleigh number:
⎧ L 1 3 m 2 L2 ⎫ ⎪ 2 TL 6 2 ⎪⎪ L t T = {1} {Ra} = ⎪⎨ t T mL m ⎬ ⎪ ⎪ t 3 T Lt ⎪⎭ ⎩⎪
(2)
We take the ratio of Ra and Gr: g β ΔT L3 ρ 2 c p
Ratio of Rayleigh number and Grashof number:
Ra kμ = g β ΔT L3 ρ 2 Gr
=
cp μ k
(3)
μ2 We recognize Eq. 3 as the Prandtl number, Prandtl number:
Pr =
Ra c p μ ρ c p μ ν = = = Gr ρk α k
(4)
Discussion Many of the established nondimensional parameters are formed by the ratio or product of two (or more) other established nondimensional parameters.
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Chapter 7 Dimensional Analysis and Modeling 7-48 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are five parameters in this problem; n = 5, h = f (ω , ρ , g , R )
List of relevant parameters:
n=5
(1)
Step 2 The primary dimensions of each parameter are listed,
h
ω
ρ
g
R
{L }
{t }
{m L }
{L t }
{L }
−1
1
1 −3
1 −2
1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j=3
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 5−3 = 2
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we elect not to pick the viscosity. We choose
ω, ρ, and R
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = hω a1 ρ b1 R c1
{( L )( t 1
−1
) (m L ) (L ) a1
1 −3
b1
1
c1
}
mass:
{m } = {m }
0 = b1
b1 = 0
time:
{t } = {t }
0 = − a1
a1 = 0
0 = 1 − 3b1 + c1
c1 = −1
b1
0
− a1
0
{L } = { L L
length:
1 −3b1
0
Lc1
}
The dependent Π is thus Π1:
Π1 =
h R
The second Pi (the only independent Π in this problem) is generated:
{Π 2 } =
Π 2 = gω a2 ρ b2 R c2 mass:
{m } = {m }
time:
{ t } = {t
b2
0
0
−2 − a2
t
}
{( L t
1 −2
)( t ) ( m L ) ( L ) −1
a2
1 −3
b2
1
c2
}
0 = b2
b2 = 0
0 = −2 − a2
a2 = −2
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Chapter 7 Dimensional Analysis and Modeling
length:
{L } = {L L 0
1 −3b2
Lc2
}
c2 = −1
0 = 1 − 3b2 + c2 0 = 1 + c2
which yields
Π2:
Π2 =
g ω2R
If we take Π2 to the power –1/2 and recognize that ωR is the speed of the rim, we see that Π2 can be modified into a Froude number,
Modified Π2:
Π 2 = Fr =
ωR gR
Step 6 We write the final functional relationship as
Relationship between Πs:
h = f ( Fr ) R
(2)
Discussion In the generation of the first Π, h and R have the same dimensions. Thus, we could have immediately written down the result, Π1 = h/R. Notice that density ρ does not appear in the result. Thus, density is not a relevant parameter after all.
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Chapter 7 Dimensional Analysis and Modeling 7-49 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are seven parameters in this problem; n = 7, h = f ( ω , ρ , g , R, t , μ )
List of relevant parameters:
n=7
(1)
Step 2 The primary dimensions of each parameter are listed,
h
ω
ρ
g
R
t
μ
{L }
{t }
{m L }
{L t }
{L }
{t }
{m L t }
−1
1
1 −3
1 −2
1
1
1 −1 −1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j =3
Reduction: If this value of j is correct, the expected number of Πs is
Number of expected Πs:
k = n− j = 7−3 = 4
Step 4 We need to choose three repeating parameters since j = 3. For convenience we choose the same repeating parameters that we used in the previous problem,
ω, ρ, and R
Repeating parameters:
Step 5 The first two Πs are identical to those of the previous problem:
Π1:
Π1 =
h R
and
Π2:
Π2 =
ωR gR
where Π2 is identified as a form of the Froude number. The third Π is formed with time t. Since repeating parameter ω has dimensions of 1/time, it is the only one that remains in the Π. Thus, without the formal algebra, Π 3 = ωt
Π3: Finally, Π4 is generated with liquid viscosity,
{Π 4 } =
Π 4 = μω a4 ρ b4 R c4 mass:
{m } = {m m }
time:
{t } = { t
length:
0
1
0
{L } = { L 0
b4
−1 − a4
t
−1 −3b1
L
}
Lc1
}
{( m L t
1 −1 −1
)( t ) ( m L ) ( L ) a4
−1
1 −3
b4
1
c4
}
0 = 1 + b4
b4 = −1
0 = −1 − a4
a4 = −1
0 = −1 − 3b4 + c4
c4 = −2
c4 = 1 + 3b4
The final Π is thus
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Chapter 7 Dimensional Analysis and Modeling
Π4:
Π4 =
μ ρΩR 2
(2)
If we invert Π4 and recognize that ωR is the speed of the rim, it becomes clear that Π4 of Eq. 2 can be modified into a Reynolds number,
Modified Π4:
ρω R 2 = Re μ
(3)
h = f ( Fr, ωt , Re ) R
(4)
Π4 =
Step 6 We write the final functional relationship as
Relationship between Πs:
Discussion Notice that this time density ρ does appear in the result. There are other acceptable answers, but this one has the most established dimensionless groups.
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Chapter 7 Dimensional Analysis and Modeling 7-50 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are five parameters in this problem; n = 5, f k = f (V , ρ , μ , D )
List of relevant parameters:
n=5
Step 2 The primary dimensions of each parameter are listed,
fk
V
ρ
μ
D
{t }
{L t }
{m L }
{m L t }
{L }
−1
1 −1
1 −3
1 −1 −1
1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j =3
Reduction: If this value of j is correct, the expected number of Πs is
Number of expected Πs:
k = n− j = 5−3 = 2
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we elect not to pick the viscosity. We choose
V, ρ, and D
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = f kV a1 ρ b1 D c1
mass:
{m } = {m }
time:
{t } = { t
−1
)( L t ) ( m L ) ( L ) 1 −1
t
}
{L } = {L L
Lc1
0
−1 − a1
a1
0
−3b1
}
a1
1 −3
b1
1
c1
}
0 = b1
b1 = 0
0 = −1 − a1
a1 = −1
0 = a1 − 3b1 + c1
c1 = 1
b1
0
length:
{( t
The dependent Π is thus
Π1:
Π1 =
fk D = St V
where we have identified this Pi as the Strouhal number. The second Pi (the only independent Π in this problem) is generated:
{Π 2 } =
Π 2 = μV a2 ρ b2 D c2
mass:
{m } = {m m }
time:
{t } = { t
0
0
1
b2
−1 − a2
t
}
{( m L t
1 −1 −1
)( L t ) ( m L ) ( L ) 1 −1
a2
1 −3
b2
1
c2
}
0 = 1 + b2
b2 = −1
0 = −1 − a2
a2 = −1
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Chapter 7 Dimensional Analysis and Modeling
length:
{L } = {L 0
L L−3b2 Lc2
−1 a2
}
c2 = −1
0 = −1 + a2 − 3b2 + c2 0 = −1 − 1 + 3 + c2
which yields
Π2:
Π2 =
μ ρVD
We recognize this Π as the inverse of the Reynolds number. So, after inverting,
Modified Π2:
Π2 =
ρVD = Reynolds number = Re μ
Step 6 We write the final functional relationship as
Relationship between Πs:
St = f ( Re )
Discussion We cannot tell from dimensional analysis the exact form of the functional relationship. However, experiments verify that the Strouhal number is indeed a function of Reynolds number.
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Chapter 7 Dimensional Analysis and Modeling 7-51 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are six parameters in this problem; n = 6, f k = f (V , ρ , μ , D, c )
List of relevant parameters:
n=6
(1)
Step 2 The primary dimensions of each parameter are listed,
fk
V
ρ
μ
D
c
{t }
{L t }
{m L }
{m L t }
{L }
{L t }
−1
1 −1
1 −3
1 −1 −1
1
1 −1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j =3
Reduction: If this value of j is correct, the expected number of Πs is
Number of expected Πs:
k = n− j = 6−3 = 3
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we elect not to pick the viscosity. We choose
V, ρ, and D
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = f kV a1 ρ b1 D c1
mass:
{m } = {m }
time:
{t } = { t
−1
)( L t ) ( m L ) ( L ) 1 −1
t
}
{L } = {L L
Lc1
0
−1 − a1
a1
0
−3b1
}
a1
1 −3
b1
1
c1
}
0 = b1
b1 = 0
0 = −1 − a1
a1 = −1
0 = a1 − 3b1 + c1
c1 = 1
b1
0
length:
{( t
The dependent Π is thus
Π1:
Π1 =
fk D = St V
where we have identified this Pi as the Strouhal number. The second Pi (the first independent Π in this problem) is generated:
{Π 2 } =
Π 2 = μV a2 ρ b2 D c2
mass:
{m } = {m m }
time:
{t } = { t
0
0
1
b2
−1 − a2
t
}
{( m L t
1 −1 −1
)( L t ) ( m L ) ( L ) 1 −1
a2
1 −3
b2
1
c2
}
0 = 1 + b2
b2 = −1
0 = −1 − a2
a2 = −1
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Chapter 7 Dimensional Analysis and Modeling
length:
{L } = {L
L L−3b2 Lc2
−1 a2
0
}
c2 = −1
0 = −1 + a2 − 3b2 + c2 0 = −1 − 1 + 3 + c2
which yields
Π2:
Π2 =
μ ρVD
We recognize this Π as the inverse of the Reynolds number. So, after inverting,
Modified Π2:
Π2 =
ρVD = Reynolds number = Re μ
The third Pi (the second independent Π in this problem) is generated:
{Π 3 } =
Π 3 = cV a3 ρ b3 D c3
mass:
{m } = {m }
time:
{t } = { t
length:
b3
0
0
{L } = {L L 0
−1 − a3
1 a3
t
}
L−3b3 Lc3
}
{( L t
1 −1
)( L t ) ( m L ) ( L ) 1 −1
a3
1 −3
b3
1
c3
}
0 = b3
b3 = 0
0 = −1 − a3
a3 = −1
0 = 1 + a3 − 3b3 + c3
c3 = 0
0 = 1 − 1 + c3
which yields
Π3:
Π3 =
c V
We recognize this Π as the inverse of the Mach number. So, after inverting,
Modified Π3:
Π3 =
V = Mach number = Ma c
Step 6 We write the final functional relationship as
Relationship between Πs:
St = f ( Re, Ma )
Discussion We have shown all the details. After you become comfortable with the method of repeating variables, you can do some of the algebra in your head.
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Chapter 7 Dimensional Analysis and Modeling 7-52 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are five parameters in this problem; n = 5, W = f (ω , ρ , μ , D )
List of relevant parameters:
n=5
(1)
Step 2 The primary dimensions of each parameter are listed, W
ω
ρ
μ
D
{m L t }
{t }
{m L }
{m L t }
{L }
1 2 −3
−1
1 −3
1 −1 −1
1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j =3
Reduction: If this value of j is correct, the expected number of Πs is
Number of expected Πs:
k = n− j = 5−3 = 2
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we elect not to pick the viscosity. We choose
ω, ρ, and D
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = W ω a1 ρ b1 D c1
mass:
{m } = {m m }
time:
{t } = { t
0
1
0
−3 − a1
t
{L } = {L L
length:
0
b1
2
−3b1
}
Lc1
}
{( m L t
1 2 −3
)( t ) ( m L ) ( L ) −1
a1
1 −3
b1
1
c1
}
0 = 1 + b1
b1 = −1
0 = −3 − a1
a1 = −3
0 = 2 − 3b1 + c1
c1 = −5
The dependent Π is thus
Π1:
Π1 =
W = NP ρ D5ω 3
where we have defined this Pi as the power number (Table 7-5). The second Pi (the only independent Π in this problem) is generated:
{Π 2 } =
Π 2 = μω a2 ρ b2 D c2
mass:
{m } = {m m } 0
1
b2
{( m L t
1 −1 −1
)( t ) ( m L ) ( L ) −1
a2
1 −3
b2
1
c2
}
b2 = −1
0 = 1 + b2
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Chapter 7 Dimensional Analysis and Modeling
{t } = { t
time: length:
0
{L } = {L 0
−1 − a2
t
−1 −3b2
L
}
Lc2
}
0 = −1 − a2
a2 = −1
0 = −1 − 3b2 + c2
c2 = −2
0 = −1 + 3 + c2
which yields
Π2:
Π2 =
μ ρ D 2ω
Since Dω is the speed of the tip of the rotating stirrer blade, we recognize this Π as the inverse of a Reynolds number. So, after inverting,
Modified Π2:
Π2 =
ρ D 2ω ρ ( Dω ) D = = Reynolds number = Re μ μ
Step 6 We write the final functional relationship as N P = f ( Re )
Relationship between Πs:
(2)
Discussion After some practice you should be able to do some of the algebra with the exponents in your head. Also, we usually expect a type of Reynolds number when we combine viscosity with a density, a length, and some kind of speed, be it angular speed or linear speed.
7-53 Solution
We are to determine the dimensionless relationship between the given parameters
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis The dimensional analysis is identical to the previous problem except that we add two additional independent parameters, both of which have dimensions of length. The two Πs of the previous problem remain. We get two additional Πs since n is now equal to 7 instead of 5. There is no need to go through all the algebra for the two additional Πs – since their dimensions match those of one of the repeating variables (D), we know that all the exponents in the Π will be zero except the exponent for D, which will be –1. The two additional Πs are
Π3 and Π4:
Π3 =
Dtank D
Π4 =
hsurface D
The final functional relationship is
Relationship between Πs:
D h ⎛ ⎞ N P = f ⎜ Re, tank , surface ⎟ D D ⎠ ⎝
(1)
Discussion We could also manipulate our Πs so that we have other length ratios like hsurface/Dtank, etc. Any such combination is acceptable.
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Chapter 7 Dimensional Analysis and Modeling 7-54 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are five parameters in this problem; n = 5,
δ = f ( x, V , ρ , μ )
List of relevant parameters:
n=5
(1)
Step 2 The primary dimensions of each parameter are listed,
δ
x
{L }
{L }
1
ρ
V
μ
{L t } { m L } {m L t } 1 −1
1
1 −3
1 −1 −1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j=3
Reduction: If this value of j is correct, the expected number of Πs is
Number of expected Πs:
k = n− j = 5−3 = 2
Step 4 We need to choose three repeating parameters since j = 3. We pick length scale x, density ρ, and freestream velocity V. x, ρ , and V
Repeating parameters:
Step 5 The Πs are generated. Note that for the first Π we can do the algebra in our heads since the relationship is very simple. Namely, the dimensions of δ are identical to those of one of the repeating variables (x). In such a case we know that all the exponents in the Π group are zero except the one for x, which is –1. The dependent Π is thus
Π1:
Π1 =
δ x
The second Π is formed with viscosity,
{Π 2 } =
Π 2 = μ x a ρ bV c
mass:
{m } = { m m }
time:
{t } = { t
length:
{L } = { L 0
1 −1 −1
)( L ) ( m L ) ( L t ) } 1
a
1 −3
b
1 −1
c
1
b
0 = 1+ b
b = −1
−1 − c
}
0 = −1 − c
c = −1
0 = −1 + a − 3b + c 0 = −1 + a + 3 − 1
a = −1
0
0
{( m L t
−1 a
t
L L−3b Lc
}
which yields Π2:
Π2 =
μ ρVx
We recognize this Π as the inverse of the Reynolds number,
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Chapter 7 Dimensional Analysis and Modeling
Modified Π2 = Reynolds number based on x:
Π 2 = Re x
ρVx μ
Step 6 We write the final functional relationship as
δ
Relationship between Πs:
x
= f ( Re x )
Discussion We cannot determine the form of the relationship by purely dimensional reasoning since there are two Πs. However, in Chap. 10 we shall see that for a laminar boundary layer, Π1 is proportional to the square root of Π2.
7-55 Solution
We are to create a scale for volume flow rate and then define an appropriate Richardson number.
Analysis By “back of the envelope” reasoning (or by inspection), we define a volume flow rate scale as L2V. Then the Richardson number can be defined as
Richardson number:
Ri =
L5 g Δρ L5 g Δρ Lg Δρ = = 2 2 2 ρV 2 ρV ρ LV
(
)
(1)
Discussion It is perhaps more clear from the form of Eq. 1 that Richardson number is a ratio of buoyancy forces to inertial forces.
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Chapter 7 Dimensional Analysis and Modeling 7-56 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are six parameters in this problem; n = 6, u = f ( μ , V , h, ρ , y )
List of relevant parameters:
n=6
(1)
Step 2 The primary dimensions of each parameter are listed,
u
μ
V
h
ρ
y
{L t }
{m L t }
{L t }
{L }
{m L }
{L }
1 −1
1 −1 −1
1 −1
1 −3
1
1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j=3
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 6−3 = 3
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we elect not to pick the viscosity. It is better to pick a fixed length (h) rather than a variable length (y); otherwise y would appear in each Pi, which would not be desirable. We choose
V, ρ, and h
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = uV a1 ρ b1 h c1 mass:
{m } = {m }
time:
{t } = { t
length:
1 −1
0
−1 − a1
t
{L } = {L L L 1 a1
}
−3b1
Lc1
}
)( L t ) ( m L ) ( L ) 1 −1
a1
1 −3
b1
1
c1
}
0 = b1
b1 = 0
0 = −1 − a1
a1 = −1
0 = 1 + a1 − 3b1 + c1
c1 = 0
b1
0
0
{( L t
The dependent Π is thus Π1 =
Π1:
u V
The second Pi (the first independent Π in this problem) is generated:
{Π 2 } =
Π 2 = μV a2 ρ b2 h c2 mass:
{m } = {m m } 0
1
b2
{( m L t
1 −1 −1
)( L t ) ( m L ) ( L ) 1 −1
a2
1 −3
b2
1
c2
}
b2 = −1
0 = 1 + b2
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Chapter 7 Dimensional Analysis and Modeling
{t } = { t
time: length:
0
{L } = {L
−1 − a2
t
}
L L−3b2 Lc2
−1 a2
0
}
0 = −1 − a2
a2 = −1
0 = −1 + a2 − 3b2 + c2
c2 = −1
0 = −1 − 1 + 3 + c2
which yields Π2:
Π2 =
μ ρVh
We recognize this Π as the inverse of the Reynolds number. So, after inverting, Modified Π2:
Π2 =
ρVh = Reynolds number = Re μ
The third Pi (the second independent Π in this problem) is generated:
{Π 3 } =
Π 3 = yV a3 ρ b3 hc3
{( L )( L t 1
1 −1
) (m L ) (L ) a3
1 −3
b3
1
c3
}
mass:
{m } = { m }
0 = b3
b3 = 0
time:
{t } = {t }
0 = − a3
a3 = 0
0 = 1 + a3 − 3b3 + c3
c3 = −1
length:
b3
0
− a3
0
{L } = {L L 0
1 a3
L−3b3 Lc3
}
0 = 1 + c3
which yields Π3:
Π3 =
y h
Step 6 We write the final functional relationship as
Relationship between Πs:
u y⎞ ⎛ = f ⎜ Re, ⎟ V h ⎝ ⎠
(2)
Discussion We notice in the first and third Πs that when the parameter on which we are working has the same dimensions as one of the repeating parameters, the Π is simply the ratio of those two parameters (here u/V and y/h).
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Chapter 7 Dimensional Analysis and Modeling 7-57 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are seven parameters in this problem; n = 7, u = f ( μ , V , h, ρ , y, t )
List of relevant parameters:
n=7
(1)
Step 2 The primary dimensions of each parameter are listed,
u
μ
V
h
ρ
y
t
{L t }
{m L t }
{L t }
{L }
{m L }
{L }
{t }
1 −1
1 −1 −1
1 −1
1 −3
1
1
1
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j=3
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 7−3 = 4
Step 4 We need to choose three repeating parameters since j = 3. Following the guidelines outlined in this chapter, we do not pick the viscosity. It is better to pick a fixed length (h) rather than a variable length (y); otherwise y would appear in each Pi, which would not be desirable. It would also not be wise to have time appear in each parameter. We choose
V, ρ, and h
Repeating parameters:
Step 5 The Πs are generated. The first three Πs are identical to those of the previous problem, so we do not include the details here. The fourth Π is formed by joining the new parameter t to the repeating variables,
{Π 4 } =
Π 4 = tV a4 ρ b4 h c4
{( t )( L t 1
1 −1
) (m L ) (L ) a4
1 −3
b4
1
c4
}
mass:
{m } = {m }
0 = b4
b4 = 0
time:
{t } = {t t }
0 = 1 − a4
a4 = 1
0 = a4 − 3b4 + c4
c4 = −1
length:
b4
0
1 − a4
0
{L } = {L L 0
a1
−3b4
Lc4
}
This Π is thus Π4 =
Π4:
tV h
Step 6 Combining this result with the first three Πs from the previous problem,
Relationship between Πs:
Discussion
u y tV ⎞ ⎛ = f ⎜ Re, , ⎟ V h h ⎠ ⎝
(2)
As t → ∞, Π4 becomes irrelevant and the result degenerates into that of the previous problem.
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Chapter 7 Dimensional Analysis and Modeling 7-58 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are four parameters in this problem; n = 4,
(
c = f k , T , Rgas
List of relevant parameters:
)
n=4
(1)
Step 2 The primary dimensions of each parameter are listed; the ratio of specific heats k is dimensionless.
c
k
T
{L t }
{1}
{T }
1 −1
Rgas
{L t
2 −2
1
T −1
}
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (T, L, and t). j=3
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 4−3 =1
Thus we expect only one Π. Step 4 We need to choose three repeating parameters since j = 3. We only have one choice in this problem, since there are only three independent parameters on the right-hand side of Eq. 1. However, one of these is already dimensionless, so it is a Π all by itself. In this situation we reduce j by one and continue, j = 3 −1 = 2
Reduction:
If this revised value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 4−2 = 2
We now expect two Πs. We choose two repeating parameters since j = 2, Repeating parameters:
T and Rgas
Step 5 The dependent Π is generated:
{Π1} =
Π1 = cT a1 Rgas b1
time:
{ t } = {t
temperature: length:
}
{( L t
1 −1
)( T ) ( L t 1
a1
2 −2
T −1
)
b1
}
0 = −1 − 2b1
b1 = −1/ 2
{T } = {T }
a1 = b1
a1 = −1/ 2
{L } = {L L }
0 = 1 + 2b1
b1 = −1/ 2
0
0
0
−1− 2 b1
a1 − b1
1 2 b1
Fortunately the two results for exponent b1 agree. The dependent Π is thus Π1:
Π1 =
c RgasT
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Chapter 7 Dimensional Analysis and Modeling
The independent Π is already known, Π2:
Π2 = k
Step 6 We write the final functional relationship as
Relationship between Πs:
c RgasT
= f (k )
(2)
Discussion We cannot tell from dimensional analysis the exact form of the functional relationship. However, in this case the result agrees with the known equation for speed of sound in an ideal gas, c = kRgasT .
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Chapter 7 Dimensional Analysis and Modeling 7-59 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are five parameters in this problem; n = 5, c = f ( k , T , Ru , M )
List of relevant parameters:
n=5
(1)
Step 2 The primary dimensions of each parameter are listed; the ratio of specific heats k is dimensionless.
c
k
T
{L t }
{1}
{T }
1 −1
Ru
{m L t
1 2 −2
1
M
T −1 N −1
}
{m N } 1
−1
Step 3 As a first guess, j is set equal to 5, the number of primary dimensions represented in the problem (m, T, L, N, and t). j =5
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 5−5 = 0
Obviously we cannot have zero Πs. We check that we have not missed a relevant parameter. Convinced that we have included all the relevant parameters we reduce j by 1: j = 5 −1 = 4
Reduction: If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 5−4 =1
Step 4 We need to choose four repeating parameters since j = 4. We only have one choice in this problem, since there are only four independent parameters on the right-hand side of Eq. 1. However, one of these is already dimensionless, so it is a Π all by itself. In this situation we reduce j by one (again) and continue, j = 4 −1 = 3
Reduction:
If this revised value of j is correct, the expected number of Πs is Number of expected Πs:
k = n− j = 5−3 = 2
We now expect two Πs. Since j = 3 we choose three repeating parameters, Repeating parameters:
T, M, and Ru
Step 5 The dependent Π is generated:
{Π1} =
Π1 = cT a1 M b1 Ru c1
time:
{t } = {t
mass:
{m } = {m
0
0
−1 −2 c1
t
b1
}
m c1
}
{( L t
1 −1
)( T ) ( m n ) ( m L t 1
a1
1
−1
b1
0 = −1 − 2c1 0 = b1 + c1
b1 = −c1
1 2 −2
T −1 N −1
)
c1
}
c1 = −1/ 2 b1 = 1/ 2
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Chapter 7 Dimensional Analysis and Modeling
amount of matter: temperature: length:
{N } = {N 0
− b1
{T } = {T 0
a1
N − c1
T − c1
}
}
{L } = {L L } 0
1 2 c1
0 = −b1 − c1
b1 = −c1
b1 = 1/ 2
0 = a1 − c1
a1 = c1
a1 = −1/ 2 c1 = −1/ 2
0 = 1 + 2c1
Fortunately the two results for exponent b1 agree, and the two results for exponent c1 agree. (If they did not agree, we would search for algebra mistakes. Finding none we would suspect that j is not correct or that we are missing a relevant parameter in the problem.) The dependent Π is thus Π1:
Π1 =
c M RuT
The independent Π is already known, Π2:
Π2 = k
Step 6 We write the final functional relationship as Relationship between Πs:
Discussion
Π1 =
c M RuT
= f (k )
(2)
Since we know that Rgas = Ru/M, we see that the result here is the same as that of the previous problem.
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Chapter 7 Dimensional Analysis and Modeling 7-60 Solution
We are to use dimensional analysis to find the functional relationship between the given parameters.
Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are three parameters in this problem; n = 3,
(
c = f T , Rgas
List of relevant parameters:
)
n=3
(1)
Step 2 The primary dimensions of each parameter are listed,
c
T
{L t }
{T }
1 −1
Rgas
{L t
2 −2
1
T −1
}
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (T, L, and t). j=3
Reduction:
If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n − j = 3−3 = 0
Obviously this is not correct, so we re-examine our initial assumptions. We can add another variable, k (the ratio of specific heats) to our List of relevant parameters. This problem would then be identical to Problem 7-58. Instead, for instructional purposes we reduce j by one and continue, j = 3 −1 = 2
Reduction:
If this revised value of j is correct, the expected number of Πs is Number of expected Πs:
k = n − j = 3− 2 =1
We now expect only one Π. Step 4 We need to choose two repeating parameters since j = 2. We only have one choice in this problem, since there are only two independent parameters on the right-hand side of Eq. 1,
Repeating parameters:
T and Rgas
Step 5 The dependent Π is generated:
{Π1} =
Π1 = cT a1 Rgas b1
time:
{ t } = {t
temperature: length:
}
{( L t
1 −1
)( T ) ( L t 1
a1
2 −2
T −1
)
b1
}
0 = −1 − 2b1
b1 = −1/ 2
{T } = {T }
a1 = b1
a1 = −1/ 2
{L } = {L L }
0 = 1 + 2b1
b1 = −1/ 2
0
0
0
−1− 2 b1
a1 − b1
1 2 b1
Fortunately the two results for exponent b1 agree. The dependent Π is thus
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Chapter 7 Dimensional Analysis and Modeling
Π1:
Π1 =
c RgasT
Step 6 Since there is only one Π, it is a function of nothing. This is only possible if we set the Π equal to a constant. We write the final functional relationship as
Relationship between Πs:
Π1 =
c RgasT
= constant
(2)
Discussion Our result represents an interesting case of “luck”. Although we failed to include the ratio of specific heats k in our analysis, we nevertheless obtain the correct result. In fact, if we set the constant in Eq. 2 as the square root of k, our result agrees with the known equation for speed of sound in an ideal gas, c = kRgasT .
7-48 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling 7-61 Solution We are to use dimensional analysis to find the functional relationship between the given parameters, and compare to the known equation for an ideal gas. Assumptions
1 The given parameters are the only relevant ones in the problem.
Analysis Πs).
The step-by-step method of repeating variables is employed to obtain the nondimensional parameters (the
Step 1 There are three parameters in this problem; n = 3, c = f ( P, ρ )
List of relevant parameters:
n=3
(1)
Step 2 The primary dimensions of each parameter are listed,
c
P
ρ
{L t }
{m L t }
{m L }
1 −1
1 −1 −2
1 −3
Step 3 As a first guess, j is set equal to 3, the number of primary dimensions represented in the problem (m, L, and t). j =3
Reduction:
If this value of j is correct, the expected number of Πs is Number of expected Πs:
k = n − j = 3−3 = 0
Obviously this is not correct, so we re-examine our initial assumptions. If we are convinced that c is a function of only P and ρ, we reduce j by one and continue, j = 3 −1 = 2
Reduction:
If this revised value of j is correct, the expected number of Πs is Number of expected Πs:
k = n − j = 3− 2 =1
We now expect only one Π. Step 4 We need to choose two repeating parameters since j = 2. We only have one choice in this problem, since there are only two independent parameters on the right-hand side of Eq. 1,
P and ρ
Repeating parameters: Step 5 The dependent Π is generated:
{Π1} =
Π1 = cP a1 ρ b1 time:
{t } = {t
mass:
{m } = {m
length:
0
−1 −2 a1
0
{L } = {L L 0
t
a1
}
mb1
1 −2 a1
}
L−3b1
}
{( L t
1 −1
)( m L
1 −1 −2
t
) (m L ) a1
1 −3
b1
}
0 = −1 − 2a1
a1 = −1/ 2
0 = a1 + b1
b1 = 1/ 2
0 = 1 − a1 − 3b1
0=0
0 = 1+ − 1 2
3 2
Fortunately the exponents for length agree with those of mass and time. The dependent Π is thus
7-49 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling
Π1:
Π1 = c
ρ P
Step 6 Since there is only one Π, it is a function of nothing. This is only possible if we set the Π equal to a constant. We write the final functional relationship as
Relationship between Πs:
Π1 = c
ρ P
= constant, or c = constant
P
ρ
(2)
The ideal gas equation is P = ρRgasT, or P/ρ = RgasT. Thus, Eq. 2 can be written as Alternative result using ideal gas law:
c = constant RgasT
(3)
Equation 3 is indeed consistent with the equation c = kRgasT . Discussion There is no way to obtain the value of the constant in Eq. 2 or 3 solely by dimensional analysis, but it turns out that the constant is the square root of k.
7-50 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 7 Dimensional Analysis and Modeling 7-62 Solution
We are to use dimensional analysis to find a functional relationship between FD and variables V, L, and μ.
Assumptions 1 We assume Re 4000, and thus the turbulent flow assumption is verified. The diameter can also be determined directly from the third Swamee-Jain formula to be ⎡ ⎛ LV 2 D = 0.66 ⎢ε 1.25 ⎜⎜ ⎢ ⎝ gh L ⎣
⎞ ⎟ ⎟ ⎠
4.75
⎛ L + νV 9.4 ⎜⎜ ⎝ gh L
⎞ ⎟⎟ ⎠
5.2 ⎤ 0.04
⎥ ⎥ ⎦
5.2 ⎡ ⎛ ⎞ ⎤ 400 ft ⎟ ⎥ = 0.66 ⎢0 + (0.180 × 10 −3 ft 2 / s )(12 ft 3 / s ) 9.4 ⎜⎜ 2 ⎟ ⎢ ⎝ (32.2 ft/s )(50 ft) ⎠ ⎥⎦ ⎣ = 0.89 ft
0.04
Note that the difference between the two results is less than 2%. Therefore, the simple Swamee-Jain relation Discussion can be used with confidence.
8-12 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-34 Solution In fully developed laminar flow in a circular pipe, the velocity at r = R/2 is measured. The velocity at the center of the pipe (r = 0) is to be determined. Assumptions
The flow is steady, laminar, and fully developed.
The velocity profile in fully developed laminar flow in a circular pipe is given by ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠ where umax is the maximum velocity which occurs at pipe center, r = 0. At r =R/2, ⎛ ( R / 2) 2 ⎞ 3u 1 ⎟ = u max ⎛⎜1 − ⎞⎟ = max u ( R / 2) = u max ⎜⎜1 − 2 ⎟ 4 ⎝ 4⎠ R ⎝ ⎠ Analysis
u(r) = umax(1-r2/R2)
R r
Solving for umax and substituting, 4u ( R / 2 ) 4 ( 6 m/s ) umax = = = 8.00 m/s 3 3 which is the velocity at the pipe center.
0
umax
Discussion The relationship used here is valid only for fully developed laminar flow. The result would be much different if the flow were turbulent.
8-35 Solution The velocity profile in fully developed laminar flow in a circular pipe is given. The average and maximum velocities as well as the flow rate are to be determined. Assumptions
The flow is steady, laminar, and fully developed.
The velocity profile in fully developed laminar flow in a circular pipe is given by ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠ The velocity profile in this case is given by Analysis
u (r ) = 4(1 − r 2 / R 2 )
u(r) = umax(1-r2/R2)
R r
Comparing the two relations above gives the maximum velocity to be umax = 4.00 m/s. Then the average velocity and volume flow rate become u 4 m/s Vavg = max = = 2.00 m/s 2 2
0
umax
V = V avg Ac = V avg (πR 2 ) = (2 m/s)[π (0.02 m) 2 ] = 0.00251 m 3 /s
Discussion A unique feature of fully developed laminar pipe flow is that the maximum velocity is exactly twice the average velocity. This is not the case for turbulent pipe flow, since the velocity profile is much fuller.
8-13 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-36 Solution The velocity profile in fully developed laminar flow in a circular pipe is given. The average and maximum velocities as well as the flow rate are to be determined. Assumptions Analysis
The flow is steady, laminar, and fully developed.
The velocity profile in fully developed laminar flow in a circular pipe is given by ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠
u(r) = umax(1-r2/R2)
R
The velocity profile in this case is given by u (r ) = 4(1 − r / R ) 2
2
r
Comparing the two relations above gives the maximum velocity to be umax = 4.00 m/s. Then the average velocity and volume flow rate become u 4 m/s Vavg = max = = 2.00 m/s 2 2 V = V A = V (πR 2 ) = (2 m/s)[π (0.07 m) 2 ] = 0.0308 m 3 /s avg
c
0
umax
avg
Discussion Compared to the previous problem, the average velocity remains the same since the maximum velocity (at the centerline) remains the same, but the volume flow rate increases as the diameter increases.
8-37 Solution Air enters the constant spacing between the glass cover and the plate of a solar collector. The pressure drop of air in the collector is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The roughness effects are negligible, and thus the inner surfaces are considered to be smooth, ε ≈ 0. 4 Air is an ideal gas. 5 The local atmospheric pressure is 1 atm. Properties
The properties of air at 1 atm and 45° are ρ = 1.109 kg/m3, μ = 1.941×10-5 kg/m⋅s, and ν = 1.750×10-5 m2/s.
Analysis
Mass flow rate, cross-sectional area, hydraulic diameter, average velocity, and the Reynolds number are m = ρV = (1.11 kg/m 3 )(0.15 m 3 /s ) = 0.1665 kg/s
Ac = a × b = (1 m)(0.03 m) = 0.03 m 2
Glass cover
Air 0.15 m3/s
5m 4 Ac 4(0.03 m 2 ) = = 0.05825 m p 2(1 + 0.03) m 0.15 m 3 / s V = =5m/ s V= Ac 0.03 m 2 Collector plate VD h (5 m/s)(0.05825 m) Insulation Re = = = 1.664 × 10 4 2 −5 υ 1.750 × 10 m /s Since Re is greater than 4000, the flow is turbulent. The friction factor corresponding to this Reynolds number for a smooth flow section (ε/D = 0) can be obtained from the Moody chart. But to avoid reading error, we use the Colebrook equation,
Dh =
⎛ε / D 2.51 = −2.0 log⎜ + ⎜ f ⎝ 3.7 Re f which gives f = 0.0271. Then the pressure drop becomes 1
ΔP = ΔPL = f
⎞ ⎟ → ⎟ ⎠
⎛ 2.51 ⎞⎟ = −2.0 log⎜ 0 + ⎜ f 16,640 f ⎟⎠ ⎝
1
(1.11 kg/m 3 )(5 m/s) 2 L ρV 2 5m = 0.0271 D 2 0.05825 m 2
⎛ 1N ⎜ ⎜ 1 kg ⋅ m/s 2 ⎝
⎞⎛ 1 Pa ⎞ ⎟⎜ = 32.3 Pa ⎟⎝ 1 N/m 2 ⎟⎠ ⎠
Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.0270, which is sufficiently close to 0.0271.
8-14 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-38 Solution Oil flows through a pipeline that passes through icy waters of a lake. The pumping power needed to overcome pressure losses is to be determined. Assumptions The flow is steady and incompressible. 2 The flow section considered is away from the entrance, and thus the flow is fully developed. 3 The roughness effects are negligible, and thus the inner surfaces are considered to be smooth, ε ≈ 0. Properties
The properties of oil are given to be ρ = 894 kg/m3 and μ = 2.33 kg/m⋅s.
Analysis
The volume flow rate and the Reynolds number in this case are
π (0.4 m) 2 πD 2 V = VAc = V = (0.5 m/s) = 0.0628 m 3 /s 4 4
(Icy lake, 0°C) Oil 10°C 0.5 m/s
ρVD (894 kg/m 3 )(0.5 m/s)(0.4 m) Re = = = 76.7 μ 2.33 kg/m ⋅ s
D = 0.4 m
L = 300 m
which is less than 2300. Therefore, the flow is laminar, and the friction factor is 64 64 f = = = 0.834 Re 76.7 Then the pressure drop in the pipe and the required pumping power become ΔP = ΔPL = f
300 m (894 kg/m 3 )(0.5 m/s) 2 L ρV 2 = 0.834 0.4 m 2 D 2
⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎜⎜ ⎟⎟⎜ ⎟ = 69.9 kPa 2 1000 kg m/s ⋅ ⎝ ⎠⎝ 1 kN/m ⎠
⎛ 1 kW ⎞ W pump = VΔP = (0.0628 m 3 /s)(69.9 kPa )⎜ ⎟ = 4.39 kW 3 ⎝ 1 kPa ⋅ m /s ⎠ Discussion The power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be much more than this due to pump inefficiency; the electrical power input will be even more due to motor inefficiency.
8-39 Solution Laminar flow through a square channel is considered. The change in the head loss is to be determined when the average velocity is doubled. Assumptions developed. Analysis
1 The flow remains laminar at all times. 2 The entrance effects are negligible, and thus the flow is fully
The friction factor for fully developed laminar flow in a square channel is ρVD 56.92 where Re = f = Re μ V
Then the head loss for laminar flow can be expressed as h L ,1 = f
μL L V 2 56.92 L V 2 56.92μ L V 2 = = = 28.46V ρVD D 2 g D 2g Re D 2 g ρgD 2
which shows that the head loss is proportional to the average velocity. Therefore, the head loss doubles when the average velocity is doubled. This can also be shown as hL ,2 = 28.46V2
Discussion
⎛ μL μL μL ⎞ ⎜ ⎟ = 2hL ,1 2 = 28.46( 2V ) 2 = 2⎜ 28.46V ρgD ρgD ρgD 2 ⎟⎠ ⎝
The conclusion above is also valid for laminar flow in channels of different cross-sections.
8-15 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-40 Solution Turbulent flow through a smooth pipe is considered. The change in the head loss is to be determined when the average velocity is doubled. Assumptions 1 The flow remains turbulent at all times. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The inner surface of the pipe is smooth. Analysis
The friction factor for the turbulent flow in smooth pipes is given as ρVD f = 0.184 Re −0.2 where Re =
μ
Then the head loss of the fluid for turbulent flow can be expressed as h L ,1 = f
⎛ ρVD ⎞ L V2 L V2 ⎟⎟ = 0.184 Re − 0.2 = 0.184⎜⎜ D 2g D 2g ⎝ μ ⎠
−0.2
⎛ ρD ⎞ L V2 ⎟⎟ = 0.184⎜⎜ D 2g ⎝ μ ⎠
−0.2
L V 1.8 D 2g
which shows that the head loss is proportional to the 1.8th power of the average velocity. Therefore, the head loss increases by a factor of 21.8 = 3.48 when the average velocity is doubled. This can also be shown as −0.2
−0.2
⎛ ρD ⎞ ⎛ ρD ⎞ L V21.8 L (2V )1.8 ⎟⎟ ⎟⎟ = 0.184⎜⎜ hL ,2 = 0.184⎜⎜ D 2g D 2g ⎝ μ ⎠ ⎝ μ ⎠ 0 . 2 − ⎡ ⎛ ρD ⎞ L V 1.8 ⎤ ⎟⎟ ⎥ = 21.8 hL ,1 = 3.48hL ,1 = 21.8 ⎢0.184⎜⎜ D 2g ⎥ ⎢⎣ ⎝ μ ⎠ ⎦ For fully rough flow in a rough pipe, the friction factor is independent of the Reynolds number and thus the flow velocity. Therefore, the head loss increases by a factor of 4 in this case since
V
L V2 D 2g and thus the head loss is proportional to the square of the average velocity when f, L, and D are constant. h L ,1 = f
Discussion Most flows in practice are in the fully rough regime, and thus the head loss is generally assumed to be proportional to the square of the average velocity for all kinds of turbulent flow. Note that we use diameter D here in place of hydraulic diameter Dh. For a square duct, it turns out that Dh = D, so this is a valid approximation.
8-16 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-41 Solution
Air enters a rectangular duct. The fan power needed to overcome the pressure losses is to be determined.
Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 Air is an ideal gas. 4 The duct involves no components such as bends, valves, and connectors. 5 The flow section involves no work devices such as fans or turbines Properties The properties of air at 1 atm and 35°C are ρ = 1.145 kg/m3, μ = 1.895×10-5 kg/m⋅s, and ν = 1.655×10-5 2 m /s. The roughness of commercial steel surfaces is ε = 0.000045 m. Analysis
The hydraulic diameter, the volume flow rate, and the Reynolds number in this case are 4 Ac 4(0.15 m)(0.20 m) 4ab Dh = = = = 0.1714 m 7m p 2(a + b) 2(0.15 + 0.20) m V = VA = V (a × b) = (7 m/s)(0.15 × 0.20 m 2 ) = 0.21 m 3 /s c
Re =
ρVD h (1.145 kg/m 3 )(7 m/s)(0.1714 m) = = 72,490 μ 1.895 × 10 −5 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is
15 cm
Air 7 m/s
20 cm
4.5 × 10 −5 m ε / Dh = = 2.625 × 10 − 4 0.1714 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ 2.625 × 10 −4 ⎛ ε / Dh 2.51 ⎞⎟ 1 2.51 ⎞⎟ → = −2.0 log⎜ + = −2.0 log⎜ + ⎜ ⎟ ⎜ 3.7 f Re f ⎠ f 72,490 f ⎟⎠ ⎝ ⎝ 3.7 It gives f = 0.02034. Then the pressure drop in the duct and the required pumping power become 1
ΔP = ΔPL = f
(1.145 kg/m 3 )(7 m/s) 2 L ρV 2 7m = 0.02034 D 2 0.1714 m 2
⎛ 1N ⎜ ⎜ 1 kg ⋅ m/s 2 ⎝
⎞⎛ 1 Pa ⎞ ⎟⎜ ⎟⎝ 1 N/m 2 ⎟⎠ = 23.3 Pa ⎠
⎛ 1W ⎞ Wpump = V ΔP = ( 0.21 m3 /s ) ( 23.3 Pa ) ⎜ ⎟ = 4.90 W 3 ⎝ 1 Pa ⋅ m /s ⎠ Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.02005, which is sufficiently close to 0.02034. Also, the power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be much more than this due to fan inefficiency; the electrical power input will be even more due to motor inefficiency.
8-17 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-42E Solution Water passes through copper tubes at a specified rate. The pumping power required per ft length to maintain flow is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as pumps and turbines. Properties The density and dynamic viscosity of water at 60°F are ρ = 62.36 lbm/ft3 and μ = 2.713 lbm/ft⋅h = -4 7.536×10 lbm/ft⋅s. The roughness of copper tubing is 5×10-6 ft. Analysis
First we calculate the average velocity and the Reynolds number to determine the flow regime: m m 1.2 lbm / s = = = 6.272 ft/s V= 2 ρAc ρ (πD / 4 ) (62.36 lbm/ft 3 )[π (0.75/12 ft) 2 / 4]
ρVD (62.36 lbm/ft 3 )(6.272 ft/s)(0.75/12 ft) = = 32,440 μ 7.536 × 10 − 4 lbm/ft ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Re =
ε/D=
Water
D = 0.75 in
1.2 lbm/s
5 × 10 −6 ft = 8 × 10 −5 0.75 / 12 ft
The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D 2.51 = −2.0 log⎜ + ⎜ f ⎝ 3.7 Re f
1
⎞ ⎟ → ⎟ ⎠
⎛ 8 × 10 −5 2.51 ⎞⎟ = −2.0 log⎜ + ⎜ f 32,440 f ⎟⎠ ⎝ 3.7
1
It gives f = 0.02328. Then the pressure drop and the required power input become
ΔP = ΔPL = f
(62.36 lbm/ft 3 )(6.272 ft/s) 2 ⎛ L ρV 2 1 ft 1 lbf ⎞ = 14.2 lbf/ft 2 = 0.02328 ⎜ ⎟ D 2 0.75/12 ft 2 ⎝ 32.2 lbm ⋅ ft/s ⎠
1W m ΔP (1.2 lbm/s)(14.2 lbf/ft 2 ) ⎛ ⎞ = 0.37 W (per ft length) = W pump = VΔP = ⎜ ⎟ ρ ⎝ 0.737 lbf ⋅ ft/s ⎠ 62.36 lbm/ft 3 Therefore, useful power input in the amount of 0.37 W is needed per ft of tube length to overcome the frictional losses in the pipe. Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.02305, which is sufficiently close to 0.02328. Also, the friction factor corresponding to ε = 0 in this case is 0.02306, which indicates that copper pipes can be assumed to be smooth with a negligible error. Also, the power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be more than this due to pump inefficiency; the electrical power input will be even more due to motor inefficiency.
8-18 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-43 Solution The pressure of oil in a pipe which discharges into the atmosphere is measured at a certain location. The flow rates are to be determined for 3 different orientations. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is laminar (to be verified). 4 The pipe involves no components such as bends, valves, and connectors. 5 The piping section involves no work devices such as pumps and turbines. Properties
The density and dynamic viscosity of oil are given to be ρ = 876 kg/m3 and μ = 0.24 kg/m⋅s.
Analysis
The pressure drop across the pipe and the cross-sectional area are ΔP = P1 − P2 = 135 − 88 = 47 kPa Ac = πD 2 / 4 = π (0.015 m) 2 / 4 = 1.767 × 10 − 4 m 2
135 kPa
(a) The flow rate for all three cases can be determined from, (ΔP − ρgL sin θ )πD 4 V = 128μL where θ is the angle the pipe makes with the horizontal. For the horizontal case, θ = 0 and thus sin θ = 0. Therefore,
(47 kPa)π (0.015 m) 4 ⎛⎜ 1 kg ⋅ m/s 2 ΔPπD 4 = Vhoriz = 128μL 128(0.24 kg/m ⋅ s)(15 m) ⎜⎝ 1 N
⎞⎛ 1000 N/m 2 ⎟⎜ ⎟⎜ 1 kPa ⎠⎝
Oil
L = 15 m
D = 1.5 cm
⎞ ⎟ = 1.62 × 10 −5 m 3 /s ⎟ ⎠
(b) For uphill flow with an inclination of 8°, we have θ = +8°, and (ΔP − ρgL sin θ )πD 4 Vuphill = 128μL =
[(47,000 Pa − (876 kg/m 3 )(9.81 m/s 2 )(15 m) sin 8°]π (0.015 m) 4 128(0.24 kg/m ⋅ s)(15 m)
⎛ 1 kg ⋅ m/s 2 ⎜ ⎜ 1 Pa ⋅ m 2 ⎝
⎞ ⎟ ⎟ ⎠
= 1.00 × 10 −5 m 3 /s
(c) For downhill flow with an inclination of 8°, we have θ = -8°, and (ΔP − ρgL sin θ )πD 4 Vdownhill = 128μL =
[(47,000 Pa − (876 kg/m 3 )(9.81 m/s 2 )(15 m) sin( −8°)]π (0.015 m) 4 128(0.24 kg/m ⋅ s)(15 m)
⎛ 1 kg ⋅ m/s 2 ⎜ ⎜ 1 Pa ⋅ m 2 ⎝
⎞ ⎟ ⎟ ⎠
= 2.24 × 10 −5 m 3 /s
The flow rate is the highest for downhill flow case, as expected. The average fluid velocity and the Reynolds number in this case are V=
2.24 × 10 −5 .m 3 /s V = = 0.127 m/s Ac 1.767 × 10 - 4 m 2
Re =
ρVD (876 kg/m 3 )(0.127 m/s)(0.015 m) = = 7.0 0.24 kg/m ⋅ s μ
which is less than 2300. Therefore, the flow is laminar for all three cases, and the analysis above is valid. Discussion Note that the flow is driven by the combined effect of pressure difference and gravity. As can be seen from the calculated rates above, gravity opposes uphill flow, but helps downhill flow. Gravity has no effect on the flow rate in the horizontal case.
8-19 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-44 Solution Glycerin is flowing through a horizontal pipe which discharges into the atmosphere at a specified flow rate. The absolute pressure at a specified location in the pipe, and the angle θ that the pipe must be inclined downwards for the pressure in the entire pipe to be atmospheric pressure are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is laminar (to be verified). 4 The pipe involves no components such as bends, valves, and connectors. 5 The piping section involves no work devices such as pumps and turbines. Properties
The density and dynamic viscosity of glycerin at 40°C are given to be ρ = 1252 kg/m3 and μ = 0.27 kg/m⋅s.
Analysis
(a) The flow rate for horizontal or inclined pipe can be determined from
(ΔP − ρgL sin θ )πD 4 V = (1) 128μL where θ is the angle the pipe makes with the horizontal. For the horizontal case, θ = 0 and thus sin θ = 0. Therefore, ΔPπD 4 Vhoriz = 128μL Solving for ΔP and substituting, ΔP =
128μLVhoriz
πD
4
=
Glycerin
D = 2 cm
0.035 L/s
L = 25 m
(2)
128(0.27 kg/m ⋅ s)(25 m)(0.035 × 10 -3 m 3 /s) ⎛⎜ 1 kN 4 2 ⎜ π (0.02 m) ⎝ 1000 kg ⋅ m/s
⎞ ⎟ ⎟ ⎠
= 60.2 kN/m 2 = 60.2 kPa
Then the pressure 25 m before the pipe exit becomes ΔP = P1 − P2
→
P1 = P2 + ΔP = 100 + 60.2 = 160.2 kPa
(b) When the flow is gravity driven downhill with an inclination θ, and the pressure in the entire pipe is constant at the atmospheric pressure, the hydrostatic pressure rise with depth is equal to pressure drop along the pipe due to frictional effects. Setting ΔP = P1 – P2 = 0 in Eq. (1) and substituting, θ is determined to be
ρg sin θπD 4 Vdownhill = 128μ 0.035 × 10 −3 m 3 /s =
− (1252 kg/m 3 )(9.81 m/s 2 ) sin θπ (0.02 m) 4 128(0.27 kg/m ⋅ s)
→
θ = −11.3°
Therefore, the pipe must be inclined 11.3° downwards from the horizontal to maintain flow in the pipe at the same rate. Verification: The average fluid velocity and the Reynolds number in this case are V=
0.035 × 10 −3 m 3 /s V V = = = 0.111 m/s Ac πD 2 / 4 π (0.02 m) 2 / 4
Re =
ρVD (1252 kg/m 3 )(0.111 m/s)(0.02 m) = = 10.3 μ 0.27 kg/m ⋅ s
which is less than 2300. Therefore, the flow is laminar, as assumed, and the analysis above is valid. Discussion Note that the flow is driven by the combined effect of pressure difference and gravity. Gravity has no effect on the flow rate in the horizontal case, but it governs the flow alone when there is no pressure difference across the pipe.
8-20 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-45 Solution Air in a heating system is distributed through a rectangular duct made of commercial steel at a specified rate. The pressure drop and head loss through a section of the duct are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 Air is an ideal gas. 4 The duct involves no components such as bends, valves, and connectors. 5 The flow section involves no work devices such as fans or turbines. Properties The roughness of commercial steel surfaces is ε = 0.000045 m. The dynamic viscosity of air at 40°C is μ = 1.918×10-5 kg/m⋅s, and it is independent of pressure. The density of air listed in that table is for 1 atm. The density at 105 kPa and 315 K can be determined from the ideal gas relation to be
ρ=
P 105 kPa = = 1.169 kg/m 3 RT (0.287 kPa ⋅ m 3 /kg.K)(40 + 273 K)
Air duct 0.2 m × 0.3 m
Analysis The hydraulic diameter, average velocity, and Reynolds number are 4 Ac 4(0.3 m)(0.20 m) 4ab Dh = = = = 0.24 m p 2(a + b) 2(0.3 + 0.20) m
Re =
L = 40
Air 43°C 0.5 m3/s
0.5.m 3 /s V V V= = = = 8.333 m/s Ac a × b (0.3 m)(0.2 m)
ρVDh (1.169 kg/m 3 )(8.333 m/s)(0.24 m) = = 121,900 μ 1.918 × 10 −5 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the duct is 4.5 × 10 −5 m = 1.875 × 10 − 4 0.24 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme),
ε / Dh =
⎛ ε / Dh ⎛ 1.875 × 10 −4 ⎞ 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 ⎟ ⎜ ⎟ 3 . 7 f Re f f 121 , 900 f ⎝ ⎠ ⎝ ⎠ It gives f = 0.01833. Then the pressure drop in the duct and the head loss become 1
ΔP = ΔPL = f hL =
L ρV 2 40 m (1.169 kg/m 3 )(8.333 m/s) 2 = 0.01833 D 2 0.24 m 2
⎛ 1N ⎜ ⎜ 1 kg ⋅ m/s 2 ⎝
⎞ ⎟ = 124 N/m 2 = 124 Pa ⎟ ⎠
ΔPL L V2 40 m (8.333 m/s) 2 = f = 0.01833 = 10.8 m ρg D 2g 0.24 m 2(9.81 m/s 2 )
Discussion
The required pumping power in this case is
⎛ 1W ⎞ W pump = VΔP = (0.5 m 3 /s)(124 Pa )⎜ ⎟ = 62 W ⎝ 1 Pa ⋅ m 3 /s ⎠ Therefore, 62 W of mechanical power needs to be imparted to the fluid. The shaft power will be more than this due to fan inefficiency; the electrical power input will be even more due to motor inefficiency. Also, the friction factor could be determined easily from the explicit Haaland relation. It would give f = 0.0181, which is sufficiently close to 0.0183.
8-21 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-46 Solution Glycerin is flowing through a smooth pipe with a specified average velocity. The pressure drop per 10 m of the pipe is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as pumps and turbines. Properties respectively.
The density and dynamic viscosity of glycerin at 40°C are given to be ρ = 1252 kg/m3 and μ = 0.27 kg/m⋅s,
Analysis
The volume flow rate and the Reynolds number are V = VAc = V (πD 2 / 4) = (3.5 m/s)[π (0.05 m) 2 / 4] = 0.006872 m 3 /s Re =
ρVD h (1252 kg/m 3 )(3.5 m/s)(0.05 m) = = 811.5 μ 0.27 kg/m ⋅ s
Glycerin 3.5 m/s
which is less than 2300. Therefore, the flow is laminar, and the friction factor for this circular pipe is 64 64 f = = = 0.07887 Re 811.5 Then the pressure drop in the pipe becomes ΔP = ΔPL = f
L ρV 2 10 m (1252 kg/m 3 )(3.5 m/s) 2 = 0.07887 D 2 0.05 m 2
D = 5 cm
L = 10 m
⎞⎛ 1 kPa ⎛ 1 kN ⎟⎟⎜ ⎜⎜ 2 1000 kg m/s ⋅ ⎠⎝ 1 kN/m ⎝
⎞ ⎟ = 121 kPa ⎠
The required pumping power in this case is ⎛ 1 kW ⎞ W pump = VΔP = (0.006872 m 3 /s )(121 kPa )⎜ ⎟ = 0.83 kW ⎝ 1 kPa ⋅ m 3 /s ⎠ Therefore, 0.83 kW of mechanical power needs to be imparted to the fluid. The shaft power will be more than this due to pump inefficiency; the electrical power input will be even more due to motor inefficiency.
Discussion
8-22 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
8-47
Solution In the previous problem, the effect of the pipe diameter on the pressure drop for the same constant flow rate is to be investigated by varying the pipe diameter from 1 cm to 10 cm in increments of 1 cm. Analysis
The EES Equations window is printed below, followed by the tabulated and plotted results.
g=9.81 Vdot=3.5*pi*(0.05)^2/4 Ac=pi*D^2/4 rho= 1252 nu=mu/rho mu=0.27 L= 10 V=Vdot/Ac "Reynolds number" Re=V*D/nu f=64/Re DP=f*(L/D)*rho*V^2/2000 "kPa" W=Vdot*DP "kW"
D, m
ΔP, kPa
V, m/s
Re
0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.1
75600 4725 933.3 295.3 121 58.33 31.49 18.46 11.52 7.56
87.5 21.88 9.722 5.469 3.5 2.431 1.786 1.367 1.08 0.875
4057 2029 1352 1014 811.5 676.2 579.6 507.2 450.8 405.7
80000 70000 60000
DP, kPa
50000 40000 30000 20000 10000 0 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09
0.1
D, m Discussion The pressure drop decays quite rapidly with increasing diameter – by several orders of magnitude, in fact. We conclude that larger diameter pipes are better when pressure drop is of concern. Of course, bigger pipes cost more and take up more space, so there is typically an optimum pipe size that is a compromise between cost and practicality.
8-23 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-48E Solution Air is flowing through a square duct made of commercial steel at a specified rate. The pressure drop and head loss per ft of duct are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 Air is an ideal gas. 4 The duct involves no components such as bends, valves, and connectors. 5 The flow section involves no work devices such as fans or turbines. Properties The density and dynamic viscosity of air at 1 atm and 60°F are ρ = 0.07633 lbm/ft3, μ = 0.04365 lbm/ft⋅h, and ν = 0.5718 ft2/s = 1.588×10-4 ft2/s. The roughness of commercial steel surfaces is ε = 0.00015 ft.
The hydraulic diameter, the average velocity, and the Reynolds number in this case are
Analysis Dh =
4 Ac 4a 2 = = a = 1 ft p 4a
V V 1200 ft 3 /min V= = 2 = = 1200 ft/min = 20 ft/s Ac a (1 ft) 2 Re =
VD h
=
(20 ft/s)(1 ft)
Air duct 1 ft × 1 ft
= 1.259 × 10 5
Air ν L = 1 ft 1.588 × 10 − 4 ft 2 /s 60°F which is greater than 4000. Therefore, the flow is turbulent. 1200 cfm The relative roughness of the duct is 0.00015 ft ε / Dh = = 1.5 × 10 − 4 1 ft The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme),
⎛ ε / Dh ⎛ 1.5 × 10 −4 ⎞ 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 ⎟ ⎜ ⎟ 3 . 7 f Re f f 125 , 900 f ⎝ ⎠ ⎝ ⎠ It gives f = 0.0180. Then the pressure drop in the duct and the head loss become 1
ΔP = ΔPL = f hL =
Discussion
1 ft (0.07633 lbm/ft 3 )(20 ft/s) 2 L ρV 2 = 0.0180 1 ft 2 D 2
1 lbf ⎛ ⎞ −3 2 ⎜ ⎟ = 8.53 × 10 lbf/ft 32 . 2 lbm ⋅ ft/s ⎝ ⎠
ΔPL L V2 1 ft (20 ft/s) 2 = f = 0.0180 = 0.112 ft ρg D 2g 1 ft 2(32.2 ft/s 2 )
The required pumping power in this case is
1W ⎛ ⎞ W pump = VΔP = (1200 / 60 ft 3 /s)(8.53 ×10 −3 lbf/ft 2 )⎜ ⎟ = 0.231 W (per ft length) ⎝ 0.737 lbf ⋅ ft/s ⎠ Therefore, 0.231 W of mechanical power needs to be imparted to the fluid per ft length of the duct. The shaft power will be more than this due to fan inefficiency; the electrical power input will be even more due to motor inefficiency. Also, the friction factor could be determined easily from the explicit Haaland relation. It would give f = 0.0178, which is sufficiently close to 0.0180.
8-24 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-49 Solution Liquid ammonia is flowing through a copper tube at a specified mass flow rate. The pressure drop, the head loss, and the pumping power required to overcome the frictional losses in the tube are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The pipe involves no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as pumps and turbines. Properties The density and dynamic viscosity of liquid ammonia at -20°C are ρ = 665.1 kg/m3 and μ = 2.361×10-4 kg/m⋅s. The roughness of copper tubing is 1.5×10-6 m. Analysis
First we calculate the average velocity and the Reynolds number to determine the flow regime: 0.15 kg/s m m = = = 11.49 m/s V= 2 ρAc ρ (πD / 4 ) (665.1 kg/m 3 )[π (0.005 m) 2 / 4]
ρVD (665.1 kg/m 3 )(11.49 m/s)(0.005 m) = = 1.618 × 10 5 μ 2.361× 10 − 4 kg/m ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Re =
Ammonia
D = 5 mm
0.15 kg/s
L = 30 m
1.5 × 10 −6 m ε/D= = 3 × 10 − 4 0.005 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 3 × 10 −4 ⎞ 1 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + 5 ⎜ 3.7 Re f ⎟ ⎜ 3.7 ⎟ × f f 1 . 618 10 f ⎝ ⎠ ⎝ ⎠ It gives f = 0.01819. Then the pressure drop, the head loss, and the useful pumping power required become ΔP = ΔPL = f
L ρV 2 D 2
30 m ( 665.1 kg/m ) (11.49 m/s ) ⎛ 1 kN ⎜ 0.005 m 2 1000 kg ⋅ m/s 2 ⎝ 3
= 0.01819 hL =
2
⎞ ⎛ 1 kPa ⎟⎜ 2 ⎠ ⎝ 1 kN/m
⎞ ⎟ = 4792 kPa ≅ 4790 kPa ⎠
ΔPL L V2 30 m (11.49 m/s) 2 = f = 0.01819 = 734 m ρg D 2g 0.005 m 2(9.81 m/s 2 ) m ΔP (0.15 kg/s )(4792 kPa ) ⎛ 1 kW ⎞ = = 1.08 kW W pump = VΔP = ⎜ 3 3 ⎟ ρ 665.1 kg/m ⎝ 1 kPa ⋅ m /s ⎠
Therefore, useful power input in the amount of 1.08 kW is needed to overcome the frictional losses in the tube. Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.0180, which is sufficiently close to 0.0182. The friction factor corresponding to ε = 0 in this case is 0.0163, which is about 10% lower. Therefore, the copper tubes in this case are nearly “smooth”. Also, the power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be more than this due to pump inefficiency; the electrical power input will be even more due to motor inefficiency.
8-25 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-50 [Also solved using EES on enclosed DVD]
Solution Water is flowing through a brass tube bank of a heat exchanger at a specified flow rate. The pressure drop and the pumping power required are to be determined. Also, the percent reduction in the flow rate of water through the tubes is to be determined after scale build-up on the inner surfaces of the tubes. Assumptions 1 The flow is steady, horizontal, and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed (this is a questionable assumption since the tubes are short, and it will be verified). 3 The inlet, exit, and header losses are negligible, and the tubes involve no components such as bends, valves, and connectors. 4 The piping section involves no work devices such as pumps and turbines. Properties The density and dynamic viscosity of water at 20°C are ρ = 983.3 kg/m3 and μ = 0.467×10-3 kg/m⋅s, respectively. The roughness of brass tubing is 1.5×10-6 m. Analysis First we calculate the average velocity and the Reynolds number to determine the flow regime: 0.015 m 3 /s V V = = = 2.387 m/s V= Ac N tune (πD 2 / 4 ) 80[π (0.01 m) 2 / 4] Re =
80 tubes
ρVD h (983.3 kg/m 3 )(2.387 m/s)(0.01 m) = = 50,270 μ 0.467 × 10 −3 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The 1.5 m relative roughness of the pipe is 1.5 × 10 −6 m 1 cm ε/D= = 1.5 × 10 − 4 Water 0.01 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 1.5 × 10 −4 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 Re f ⎟ ⎜ 3 .7 f f 50,270 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.0214. Then the pressure drop, the head loss, and the useful pumping power required become ⎞⎛ 1 kPa ⎞ L ρV 2 1.5 m (983.3 kg/m 3 )(2.387 m/s) 2 ⎛ 1 kN ⎜ ⎟ ΔP = ΔPL = f = 0.0214 = 8.99 kPa 2 ⎟⎜ 2 ⎟ ⎜ D 2 0.01 m 2 ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN/m ⎠ ⎛ 1 kW ⎞ W pump = VΔP = (0.015 m 3 /s)(8.99 kPa )⎜ ⎟ = 0.135 kW 3 ⎝ 1 kPa ⋅ m /s ⎠ Therefore, useful power input in the amount of 0.135 kW is needed to overcome the frictional losses in the tube. The hydrodynamic entry length in this case is L h, turbulent ≈ 10 D = 10(0.01 m) = 0.1 m
which is much less than 1.5 m. Therefore, the assumption of fully developed flow is valid. (The effect of the entry region is to increase the friction factor, and thus the pressure drop and pumping power). After scale buildup: When 1-mm thick scale builds up on the inner surfaces (and thus the diameter is reduced to 0.8 cm from 1 cm) with an equivalent roughness of 0.4 mm, and the useful power input is fixed at 0.135 kW, the problem can be formulated as follows (note that the flow rate and thus the average velocity are unknown in this case): V=
V V = Ac N tune (πD 2 / 4)
Re =
ρVD → μ
ε/D=
Re =
→
V=
V
(1)
80[π (0.008 m) 2 / 4]
(983.3 kg/m 3 )V (0.008 m)
(2)
0.467 × 10 −3 kg/m ⋅ s
0.0004 m = 0.05 0.008 m
8-26 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes ⎛ε / D 2.51 = −2.0 log⎜ + ⎜ 3 . 7 f Re f ⎝
1
ΔP = f
L ρV 2 D 2
→ ΔP = f
W pump = 0.135 kW →
⎞ ⎟ → ⎟ ⎠
1.5 m (983.3 kg/m 3 )V 2 0.008 m 2
VΔP = 0.135
⎛ 0.05 2.51 = −2.0 log⎜ + ⎜ 3 . 7 f Re f ⎝
1
⎛ 1 kN ⎜ ⎜ 1000 kg ⋅ m/s 2 ⎝
⎞ ⎟ ⎟ ⎠
(3)
⎞⎛ 1 kPa ⎞ ⎟⎜ ⎟⎝ 1 kN/m 2 ⎟⎠ (4) ⎠
(5)
Solving this system of 5 equations in 5 unknown (f, Re, V, ΔP, and V ) using an equation solver (or a trial-and-error approach, by assuming a velocity value) gives f = 0.0723, Re = 28,870, V = 1.714 m/s, ΔP = 19.6 kPa, and V = 0.00689 m3/s = 6.89 L/s Then the percent reduction in the flow rate becomes Reduction ratio =
Vclean −Vdirty 15 − 6.89 = = 0.54 = 54% 15 V clean
Therefore, for the same pump input, the flow rate will be reduced to less than half of the original flow rate when the pipes were new and clean. Discussion The friction factor could also be determined easily from the explicit Haaland relation. Also, the power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be more than this due to pump inefficiency; the electrical power input will be even more due to motor inefficiency.
Minor Losses
8-51C Solution
We are to define minor loss and minor loss coefficient.
Analysis The head losses associated with the flow of a fluid through fittings, valves, bends, elbows, tees, inlets, exits, enlargements, contractions, etc. are called minor losses, and are expressed in terms of the minor loss coefficient as h KL = 2 L V /( 2 g ) Discussion
Basically, any irreversible loss that is not due to friction in long, straight sections of pipe is a minor loss.
8-52C Solution
We are to define equivalent length and its relationship to the minor loss coefficient.
Analysis Equivalent length is the length of a straight pipe which would give the same head loss as the minor loss component. It is related to the minor loss coefficient by D Lequiv = K L f Discussion Equivalent length is not as universal as minor loss coefficient because it depends on the roughness and Reynolds number of the equivalent straight section of pipe.
8-27 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-53C Solution
We are to discuss the effect of rounding a pipe inlet.
Analysis
The effect of rounding of a pipe inlet on the loss coefficient is (c) very significant.
Discussion In fact, the minor loss coefficient changes from 0.8 for a reentrant pipe inlet to about 0.03 for a wellrounded pipe inlet – quite a significant improvement.
8-54C Solution
We are to discuss the effect of rounding on a pipe outlet.
Analysis
The effect of rounding of a pipe exit on the loss coefficient is (a) negligible.
Discussion At any pipe outlet, all the kinetic energy is wasted, and the minor loss coefficient is equal to α, which is about 1.05 for fully developed turbulent pipe flow. Rounding of the outlet does not help.
8-55C Solution
We are to compare the minor losses of a gradual expansion and a gradual contraction.
Analysis A gradual expansion, in general, has a greater minor loss coefficient than a gradual contraction in pipe flow. This is due to the adverse pressure gradient in the boundary layer, which may lead to flow separation. Discussion Note, however, that pressure is “recovered” in a gradual expansion. In other words, the pressure rises in the direction of flow. Such a device is called a diffuser.
8-56C Solution
We are to discuss ways to reduce the head loss in a pipe flow with bends.
Analysis
Another way of reducing the head loss associated with turns is to install turning vanes inside the elbows.
Discussion There are many other possible answers, such as: reduce the inside wall roughness of the pipe, use a larger diameter pipe, shorten the length of pipe as much as possible, etc.
8-57C Solution
We are to compare two different ways to reduce the minor loss in pipe bends.
Analysis The loss coefficient is lower for flow through a 90° miter elbow with well-designed vanes (KL ≈ 0.2) than it is for flow through a smooth curved bend (KL ≈ 0.9). Therefore, using miter elbows with vanes results in a greater reduction in pumping power requirements. Discussion
Both values are for threaded elbows. The loss coefficients for flanged elbows are much lower.
8-28 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-58 Solution Water is to be withdrawn from a water reservoir by drilling a hole at the bottom surface. The flow rate of water through the hole is to be determined for the well-rounded and sharp-edged entrance cases. Assumptions 1 The flow is steady and incompressible. 2 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 3 The effect of the kinetic energy correction factor is disregarded, and thus α = 1. Analysis The loss coefficient is KL = 0.5 for the sharp-edged entrance, and KL = 0.03 for the well-rounded entrance. We take point 1 at the free surface of the reservoir and point 2 at the exit of the hole. We also take the reference level at the exit of the hole (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is zero (V1 = 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g
where the head loss is expressed as h L = K L z1 = α 2
V 22 V2 + KL 2 2g 2g
→
z1 = α 2
1
Water
Dhole= 1.5 cm
3m
2
V 22 + hL 2g
V2 . Substituting and solving for V2 gives 2g
→ 2 gz1 = V 22 (α 2 + K L )
→
V2 =
2 gz1 2 gz1 = α2 + KL 1+ K L
since α2 = 1. Note that in the special case of KL = 0, it reduces to the Toricelli equation V 2 = 2gz 1 , as expected. Then the volume flow rate becomes
πD hole V = AcV 2 = 4 2
2 gz1 1+ K L
Substituting the numerical values, the flow rate for both cases are determined to be
πDhole Well-rounded entrance: V = 4
2 gz1 π (0.015 m ) 2 = 1+ KL 4
2(9.81 m/s 2 )(3 m) = 1.34 × 10−3 m3 /s 1 + 0.03
πD hole V = 4
2 gz1 π (0.015m ) 2 = 1+ K L 4
2(9.81 m/s 2 )(3 m) = 1.11× 10 −3 m 3 /s 1 + 0.5
2
2
Sharp-edged entrance:
Discussion The flow rate in the case of frictionless flow (KL = 0) is 1.36×10-3 m3/s. Note that the frictional losses cause the flow rate to decrease by 1.5% for well-rounded entrance, and 18.5% for the sharp-edged entrance.
8-29 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-59 Solution Water is discharged from a water reservoir through a circular hole of diameter D at the side wall at a vertical distance H from the free surface. A relation for the “equivalent diameter” of the sharp-edged hole for use in frictionless flow relations is to be obtained. Assumptions 1 The flow is steady and incompressible. 2 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 3 The effect of the kinetic energy correction factor is disregarded, and thus α = 1. Analysis The loss coefficient is KL = 0.5 for the sharp-edged entrance, and KL = 0 for the “frictionless” flow. We take point 1 at the free surface of the reservoir and point 2 at the exit of the hole, which is also taken to be the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is zero (V1 = 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbin e, e + h L ρg 2g ρg 2g
where the head loss is expressed as h L = K L H =α2
V 22 V2 + KL 2 2g 2g
→
→
H = α2
V 22 + hL 2g
V2 . Substituting and solving for V2 gives 2g
2 gH = V 22 (α 2 + K L )
→
V2 =
2 gH 2 gH = 1+ K L α2 + KL
since α2 = 1. Then the volume flow rate becomes
πD V = AcV 2 = 4
2
2 gH 1+ K L
(1)
Note that in the special case of KL = 0 (frictionless flow), the velocity relation reduces to the Toricelli equation, V 2,frictionless = 2 gH . The flow rate in this case through a hole of De (equivalent diameter) is V = Ac ,equivV 2,frictionless =
2 πDequiv
2 gH 4 Setting Eqs. (1) and (2) equal to each other gives the desired relation for the equivalent diameter, 2 πDequiv πD 2 2 gH 2 gH = 4 4 1+ K L
(2) 1 H Water
D
2
which gives Dequiv =
D D = = 0.904 D ( 1 + K L )1 / 4 ( 1 + 0.5 )1 / 4
Dequiv < D
Note that the effect of frictional losses of a sharp-edged entrance is to reduce the diameter by about 10%. Discussion 2 = (0.904 D ) 2 = 0.82 D 2 . Also, noting that the flow rate is proportional to the square of the diameter, we have V ∝ Dequiv Therefore, the flow rate through a sharp-edged entrance is about 18% less compared to the frictionless entrance case.
8-30 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-60 Solution Water is discharged from a water reservoir through a circular hole of diameter D at the side wall at a vertical distance H from the free surface. A relation for the “equivalent diameter” of the slightly rounded hole for use in frictionless flow relations is to be obtained. Assumptions 1 The flow is steady and incompressible. 2 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 3 The effect of the kinetic energy correction factor is disregarded, and thus α = 1.
The loss coefficient is KL = 0.12 for the slightly rounded entrance, and KL = 0 for the “frictionless” flow. We take point 1 at the free surface of the reservoir and point 2 at the exit of the hole, which is also taken to be the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is zero (V1 = 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to Analysis
P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbin e, e + h L ρg 2g ρg 2g
where the head loss is expressed as h L = K L H =α2
V 22 V2 + KL 2 2g 2g
→
→
H = α2
V 22 + hL 2g
V2 . Substituting and solving for V2 gives 2g
2 gH = V 22 (α 2 + K L )
→
V2 =
2 gH 2 gH = 1+ K L α2 + KL
since α2 = 1. Then the volume flow rate becomes
πD V = AcV 2 = 4
2
2 gH 1+ K L
(1)
Note that in the special case of KL = 0 (frictionless flow), the velocity relation reduces to the Toricelli equation, V 2,frictionless = 2gz1 . The flow rate in this case through a hole of De (equivalent diameter) is V = Ac ,equivV 2,frictionless =
2 πDequiv
2 gH 4 Setting Eqs. (1) and (2) equal to each other gives the desired relation for the equivalent diameter, 2 πDequiv πD 2 2 gH 2 gH = 4 4 1+ K L
(2) 1 H Water
D
2
which gives Dequiv =
Dequiv < D
D D = = 0.972 D 1/ 4 (1 + KL ) ( 1 + 0.12 )1 / 4
Note that the effect of frictional losses of a slightly rounded entrance is to reduce the diameter by about 3%. Discussion 2 = (0.972 D ) 2 = 0.945D 2 . Also, noting that the flow rate is proportional to the square of the diameter, we have V ∝ Dequiv Therefore, the flow rate through a slightly rounded entrance is about 5% less compared to the frictionless entrance case.
8-31 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-61 Solution A horizontal water pipe has an abrupt expansion. The water velocity and pressure in the smaller diameter pipe are given. The pressure after the expansion and the error that would have occurred if the Bernoulli Equation had been used are to be determined. Assumptions 1 The flow is steady, horizontal, and incompressible. 2 The flow at both the inlet and the outlet is fully developed and turbulent with kinetic energy corrections factors of α1 = α2 = 1.06 (given). Properties
We take the density of water to be ρ = 1000 kg/m3.
Analysis
Noting that ρ = const. (incompressible flow), the downstream velocity of water is
A1 πD12 / 4 D12 (0.08 m) 2 V1 = V = V = (10 m/s) = 2.5 m/s 1 1 A2 πD 22 / 4 D 22 (0.16 m) 2 The loss coefficient for sudden expansion and the head loss can be calculated from m 1 = m 2 → ρV1 A1 = ρV 2 A2 → V 2 =
⎛ A K L = ⎜⎜1 − small ⎜ Alarge ⎝
2
2 2 ⎞ 2 ⎟ = ⎛⎜1 − D1 ⎞⎟ = ⎛⎜1 − 0.08 ⎜ 0.16 2 ⎟⎟ ⎜ D2 ⎟ ⎝ 2 ⎠ ⎝ ⎠
2
⎞ ⎟ = 0.5625 ⎟ ⎠
V2 (10 m/s) 2 h L = K L 1 = (0.5625) = 2.87 m 2g 2(9.81 m/s 2 )
2 1 Water
8 cm
16 cm 10 m/s
Noting that z1 = z2 and there are no pumps or turbines 300 kPa involved, the energy equation for the expansion section can be expressed in terms of heads as P1 V2 P V2 P1 V2 P V2 + α 1 1 + z1 + hpump,u = 2 + α 2 2 + z 2 + hturbine, e + h L → + α 1 1 = 2 + α 2 2 + hL ρg ρg ρg 2g 2g 2 g ρg 2g Solving for P2 and substituting, ⎧⎪α V 2 − α 2V 2 2 ⎫⎪ − gh L ⎬ P2 = P1 + ρ ⎨ 1 1 2 ⎪⎩ ⎪⎭ ⎧⎪1.06(10 m/s) 2 − 1.06(2.5 m/s) 2 ⎫⎪⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎟⎟⎜ = (300 kPa) + (1000 kg/m 3 )⎨ − (9.81 m/s 2 )(2.87 m)⎬⎜⎜ ⎟ 2 2 1000 kg m/s ⋅ ⎪⎩ ⎪⎭⎝ ⎠⎝ 1 kN/m ⎠ = 322 kPa Therefore, despite the head (and pressure) loss, the pressure increases from 300 kPa to 321 kPa after the expansion. This is due to the conversion of dynamic pressure to static pressure when the velocity is decreased. When the head loss is disregarded, the downstream pressure is determined from the Bernoulli equation to be P1 V12 P V2 P1 V12 P2 V 22 V 2 − V2 2 + + z1 = 2 + 2 + z 2 → + = + → P1 = P1 + ρ 1 ρg 2 g ρg 2 g ρg 2 g ρg 2 g 2 Substituting, ⎞⎛ 1 kPa ⎞ (10 m/s) 2 − (2.5 m/s) 2 ⎛ 1 kN ⎟⎟⎜ ⎜⎜ P2 = (300 kPa) + (1000 kg/m 3 ) ⎟ = 347 kPa 2 2 ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN/m ⎠
Therefore, the error in the Bernoulli equation is
Error = P2, Bernoulli – P2 = 347 – 322 = 25.0 kPa
Note that the use of the Bernoulli equation results in an error of (347 – 322) / 322 = 0.078 or 7.8%. Discussion It is common knowledge that higher pressure upstream is necessary to cause flow, and it may come as a surprise that the downstream pressure has increased after the abrupt expansion, despite the loss. This is because the sum of the three Bernoulli terms which comprise the total head, consisting of pressure head, velocity head, and elevation head, namely [P/ρg + ½V2/g+ z], drives the flow. With a geometric flow expansion, initially higher velocity head is converted to downstream pressure head, and this increase outweighs the non-convertible and non-recoverable head loss term.
8-32 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
Piping Systems and Pump Selection
8-62C Solution
We are to compare the flow rate and pressure drop in two pipes of different diameters in series.
Analysis For a piping system that involves two pipes of different diameters (but of identical length, material, and roughness) connected in series, (a) the flow rate through both pipes is the same and (b) the pressure drop through the smaller diameter pipe is larger. Discussion
The wall shear stress on the smaller pipe is larger, friction factor f is larger, and thus the head loss is higher.
8-63C Solution
We are to compare the flow rate and pressure drop in two pipes of different diameters in parallel.
Analysis For a piping system that involves two pipes of different diameters (but of identical length, material, and roughness) connected in parallel, (a) the flow rate through the larger diameter pipe is larger and (b) the pressure drop through both pipes is the same. Discussion Since the two pipes separate from each other but then later re-join, the pressure drop between the two junctions must be the same, regardless of which pipe segment is under consideration.
8-64C Solution
We are to compare the pressure drop of two different-length pipes in parallel.
Analysis The pressure drop through both pipes is the same since the pressure at a point has a single value, and the inlet and exits of these the pipes connected in parallel coincide. Discussion The length, diameter, roughness, and number and type of minor losses are all irrelevant – for any two pipes in parallel, both have the same pressure drop.
8-65C Solution We are to discuss whether the required pump head is equal to the elevation difference when irreversible head losses are negligible. Analysis Yes, when the head loss is negligible, the required pump head is equal to the elevation difference between the free surfaces of the two reservoirs. Discussion A pump in a piping system may: (1) raise the fluid’s elevation, and/or (2) increase the fluid’s kinetic energy, and/or (3) increase the fluid’s pressure, and/or (4) overcome irreversible losses. In this case, (2), (3), and (4) are zero or negligible; thus only (1) remains.
8-66C Solution
We are to explain how the operating point of a pipe/pump system is determined.
Analysis The pump installed in a piping system operates at the point where the system curve and the characteristic curve intersect. This point of intersection is called the operating point. Discussion
The volume flow rate “automatically” adjusts itself to reach the operating point.
8-33 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-67C Solution We are to draw a pump head versus flow rate chart and identify several parameters.
Head
Analysis The plot of the head loss versus the flow rate is called the system curve. The experimentally determined pump head and pump efficiency versus the flow rate curves are called characteristic curves. The pump installed in a piping system operates at the point where the system curve and the characteristic curve intersect. This point of intersection is called the operating point. Discussion and the .
hpump
ηpump
System (demand) curve Operating point
By matching the system (demand) curve
Supply curve
Flow rate
8-34 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-68 [Also solved using EES on enclosed DVD]
Solution The pumping power input to a piping system with two parallel pipes between two reservoirs is given. The flow rates are to be determined. Reservoir B zB=2 m
Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The elevations of the reservoirs remain constant. 4 The minor losses and the head loss in pipes other than the parallel pipes are said to be negligible. 5 The flows through both pipes are turbulent (to be verified).
Reservoir A zA=2 m
25 m 3 cm
1 2
Properties The density and dynamic viscosity of water at 20°C are ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s. Plastic pipes are smooth, and their roughness is zero, ε = 0.
5 cm Pump
Analysis This problem cannot be solved directly since the velocities (or flow rates) in the pipes are not known. Therefore, we would normally use a trial-and-error approach here. However, nowadays the equation solvers such as EES are widely available, and thus below we will simply set up the equations to be solved by an equation solver. The head supplied by the pump to the fluid is determined from (998 kg/m 3 )V (9.81 m/s 2 )hpump, u ρVgh pump, u W elect,in = → 7000 W = (1) 0.68 η pump -motor
We choose points A and B at the free surfaces of the two reservoirs. Noting that the fluid at both points is open to the atmosphere (and thus PA = PB = Patm) and that the fluid velocities at both points are zero (VA = VB =0), the energy equation for a control volume between these two points simplifies to PA V2 P V2 + α A A + z A + h pump, u = B + α B B + z B + hturbine, e + h L → h pump, u = ( z B − z A ) + h L 2g 2g ρg ρg or h pump, u = (9 − 2) + h L (2) where h L = h L ,1 = h L ,2
(3) (4)
We designate the 3-cm diameter pipe by 1 and the 5-cm diameter pipe by 2. The average velocity, Reynolds number, friction factor, and the head loss in each pipe are expressed as V1 V1 V1 = → V1 = 2 Ac,1 πD1 / 4 π (0.03m) 2 / 4 V V2 V2 V2 = 2 = → V2 = 2 Ac ,2 πD 2 / 4 π (0.05m) 2 / 4
V1 =
Re 1 =
ρV1 D1 μ
Re 2 =
ρV 2 D 2 μ
→ Re 1 =
1.002 × 10 −3 kg/m ⋅ s
→ Re 2 =
(998 kg/m 3 )V 2 (0.05 m) 1.002 × 10 −3 kg/m ⋅ s
⎛ε / D 2.51 1 = −2.0 log⎜ + ⎜ 3.7 Re1 f 1 f1 ⎝
f2
L1 V12 D1 2 g
h L ,2 = f 2
L 2 V 22 D2 2 g
→ →
h L ,1 = f 1
(7) (8)
⎛ 2.51 = −2.0 log⎜ 0 + ⎜ Re1 f 1 f1 ⎝
⎞ ⎟ (9) ⎟ ⎠
⎞ ⎛ ⎟ → 1 = −2.0 log⎜ 0 + 2.51 ⎟ ⎜ Re 2 f 2 f2 ⎠ ⎝
⎞ ⎟ (10 ⎟ ⎠
⎞ ⎟ → ⎟ ⎠
⎛ε / D 2.51 2 = −2.0 log⎜ + ⎜ 3.7 Re 2 f 2 ⎝
h L,1 = f 1
(6)
(998 kg/m 3 )V1 (0.03 m)
1
1
(5)
1
V12 25 m 0.03 m 2(9.81 m/s 2 )
h L ,2 = f 2
V 22 25 m 0.05 m 2(9.81 m/s 2 )
(11) (12)
8-35 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes V = V1 + V2
(13)
This is a system of 13 equations in 13 unknowns, and their simultaneous solution by an equation solver gives V = 0.0183 m 3 /s,
V1 = 0.0037 m 3 /s,
V2 = 0.0146 m 3 /s ,
V1 = 5.30 m/s, V2 = 7.42 m/s, h L = h L,1 = h L, 2 = 19.5 m , hpump,u = 26.5 m Re1 = 158,300,
Re2 = 369,700, f1 = 0.0164, f2 = 0.0139
Note that Re > 4000 for both pipes, and thus the assumption of turbulent flow is verified. Discussion This problem can also be solved by using an iterative approach, but it will be very time consuming. Equation solvers such as EES are invaluable for this kind of problems.
8-36 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-69E Solution The flow rate through a piping system connecting two reservoirs is given. The elevation of the source is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The elevations of the reservoirs remain constant. 4 There are no pumps or turbines in the piping system. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The roughness of cast iron pipe is ε = 0.00085 ft. Analysis The piping system involves 120 ft of 2-in diameter piping, a well-rounded entrance (KL = 0.03), 4 standard flanged elbows (KL = 0.3 each), a fully open gate valve (KL = 0.2), and a sharp-edged exit (KL = 1.0). We choose points 1 and 2 at the free surfaces of the two reservoirs. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm), the fluid velocities at both points are zero (V1 = V2 =0), the free surface of the lower reservoir is the reference level (z2 = 0), and that there is no pump or turbine (hpump,u = hturbine = 0), the energy equation for a control volume between these two points simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg 2g ρg 2g
→
z1 = h L
2 ⎛ L ⎞V + KL ⎟ where h L = h L ,total = h L,major + h L ,minor = ⎜ f ⎝ D ⎠ 2g since the diameter of the piping system is constant. The average velocity in the pipe and the Reynolds number are V V 10/60 ft 3 /s V= = = = 7.64 ft/s Ac πD 2 / 4 π (2 / 12 ft) 2 / 4
∑
ρVD (62.3 lbm/ft 3 )(7.64 ft/s)(2/12 ft) = = 60,700 μ 1.307 × 10 − 3 lbm/ft ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is
1
z1
Re =
0.00085 ft ε/D= = 0.0051 2 / 12 ft The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 0.0051 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ ⎟ ⎜ 3.7 Re f 3. 7 f f 60,700 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.0320. The sum of the loss coefficients is
∑K
L
120 ft 2 in
2 10 ft3/min
= K L ,entrance + 4 K L,elbow + K L, valve + K L,exit = 0.03 + 4 × 0.3 + 0.2 + 1.0 = 2.43
Then the total head loss and the elevation of the source become 2 2 120 ft ⎛ L ⎞V ⎛ ⎞ (7.64 ft/s) + = ⎜ (0.0320) + 2.43⎟ = 23.1 ft hL = ⎜ f KL ⎟ 2/12 ft ⎝ D ⎠ 2g ⎝ ⎠ 2(32.2 ft/s 2 )
∑
z1 = h L = 23.1 ft
Therefore, the free surface of the first reservoir must be 23.1 ft above the free surface of the lower reservoir to ensure water flow between the two reservoirs at the specified rate. Discussion Note that fL/D = 23.0 in this case, which is almost 10 folds of the total minor loss coefficient. Therefore, ignoring the sources of minor losses in this case would result in an error of about 10%.
8-37 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-70 Solution A water tank open to the atmosphere is initially filled with water. A sharp-edged orifice at the bottom drains to the atmosphere. The initial velocity from the tank and the time required to empty the tank are to be determined. Assumptions 1 The flow is uniform and incompressible. 2 The flow is turbulent so that the tabulated value of the loss coefficient can be used. 3 The effect of the kinetic energy correction factor is negligible, α = 1.
The loss coefficient is KL = 0.5 for a sharp-edged entrance.
Properties
Analysis (a) We take point 1 at the free surface of the tank, and point 2 at the exit of the orifice. We also take the reference level at the centerline of the orifice (z2 = 0), and take the positive direction of z to be upwards. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
where the head loss is expressed as h L = K L z1 = α 2
V22 V2 + KL 2 2g 2g
→
→
z1 = α 2
V 22 + hL 2g
V2 . Substituting and solving for V2 gives 2g
2 gz1 = V22 (α 2 + K L )
→
V2 =
1
2 gz1 α2 + KL
Water tank
where α2 = 1. Noting that initially z1 = 2 m, the initial velocity is determined to be V2 =
2 gz1 = 1+ K L
3m
2(9.81 m/s 2 )(2 m) = 5.11 m/s 1 + 0.5
2m 10 cm
2
The average discharge velocity through the orifice at any given time, in general, can be expressed as V2 =
2 gz 1+ K L
where z is the water height relative to the center of the orifice at that time. (b) We denote the diameter of the orifice by D, and the diameter of the tank by D0. The flow rate of water from the tank can be obtained by multiplying the discharge velocity by the orifice area,
πD V = AorificeV 2 = 4
2
2 gz 1+ K L
Then the amount of water that flows through the orifice during a differential time interval dt is dV = Vdt =
πD 2 4
2 gz dt 1+ K L
(1)
which, from conservation of mass, must be equal to the decrease in the volume of water in the tank, dV = Atank (−dz ) = −
πD02
dz (2) 4 where dz is the change in the water level in the tank during dt. (Note that dz is a negative quantity since the positive direction of z is upwards. Therefore, we used –dz to get a positive quantity for the amount of water discharged). Setting Eqs. (1) and (2) equal to each other and rearranging,
πD 2 4
πD 2 D2 2 gz dt = − 0 dz → dt = − 02 1+ K L 4 D
1+ K L dz 2 gz
→ dt = −
1 + K L −1 / 2 z dz 2g
D02 D
2
The last relation can be integrated easily since the variables are separated. Letting tf be the discharge time and integrating it from t = 0 when z = z1 to t = tf when z = 0 (completely drained tank) gives
∫
tf
dt = −
t =0
D02 D
2
1+ K L 2g
∫
0
z
−1 / 2
z = z1
dz →
tf =-
D02 D
2
− 1 +1
1+ K L z 2 2 g − 12 + 1
0
= z1
2 D02 D
2
1+ K L 1/ 2 z1 2g
8-38 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
Simplifying and substituting the values given, the draining time is determined to be tf =
D02 D
2
2 z1 (1 + K L ) (3 m) 2 = g (0.1 m) 2
2(2 m)(1 + 0.5) 9.81 m/s 2
= 704 s = 11.7 min
Discussion The effect of the loss coefficient KL on the draining time can be assessed by setting it equal to zero in the draining time relation. It gives t f ,zero loss =
D02 D2
2 z1 (3 m) 2 = g (0.1 m) 2
2(2 m) 9.81 m/s 2
= 575 s = 9.6 min
Note that the loss coefficient causes the draining time of the tank to increase by (11.7- 9.6)/11.7 = 0.18 or 18%, which is quite significant. Therefore, the loss coefficient should always be considered in draining processes.
8-39 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-71 Solution A water tank open to the atmosphere is initially filled with water. A sharp-edged orifice at the bottom drains to the atmosphere through a long pipe. The initial velocity from the tank and the time required to empty the tank are to be determined. Assumptions 1 The flow is uniform and incompressible. 2 The draining pipe is horizontal. 3 The flow is turbulent so that the tabulated value of the loss coefficient can be used. 4 The friction factor remains constant (in reality, it changes since the flow velocity and thus the Reynolds number changes). 5 The effect of the kinetic energy correction factor is negligible, so we set α = 1. Properties 0.015.
The loss coefficient is KL = 0.5 for a sharp-edged entrance. The friction factor of the pipe is given to be
Analysis (a) We take point 1 at the free surface of the tank, and point 2 at the exit of the pipe. We take the reference level at the centerline of the pipe (z2 = 0), and take the positive direction of z to be upwards. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 = α 2
V 22 + hL 2g
where ⎛ L + h L = h L ,total = h L ,major + h L ,minor = ⎜ f ⎝ D
∑K
L
2 2 ⎞V ⎛ L ⎞V =⎜ f + KL ⎟ ⎟ ⎠ 2g ⎝ D ⎠ 2g
since the diameter of the piping system is constant. Substituting and solving for V2 gives z1 = α 2
2 V 22 ⎛ L ⎞V +⎜ f + KL ⎟ 2 2g ⎝ D ⎠ 2g
→
V2 =
2 gz1 α 2 + fL / D + K L
where α2 = 1. Noting that initially z1 = 2 m, the initial velocity is determined to be V 2,i =
1 Water tank
2 gz1 2(9.81 m/s 2 )(2 m) = = 1.54 m/s 1 + fL / D + K L 1 + 0.015(100 m)/(0.1 m) + 0.5
3m
2m 10 cm
2
100 m
The average discharge velocity at any given time, in general, can be expressed as V2 =
2 gz 1 + fL / D + K L
where z is the water height relative to the center of the orifice at that time. (b) We denote the diameter of the pipe by D, and the diameter of the tank by Do. The flow rate of water from the tank can be obtained by multiplying the discharge velocity by the pipe cross-sectional area,
πD V = A pipeV 2 = 4
2
2 gz 1 + fL / D + K L
Then the amount of water that flows through the pipe during a differential time interval dt is
πD dV = Vdt = 4
2
2 gz dt 1 + fL / D + K L
(1)
which, from conservation of mass, must be equal to the decrease in the volume of water in the tank, dV = Atan k (−dz ) = −
πD02 4
dz
(2)
where dz is the change in the water level in the tank during dt. (Note that dz is a negative quantity since the positive direction of z is upwards. Therefore, we used –dz to get a positive quantity for the amount of water discharged). Setting Eqs. (1) and (2) equal to each other and rearranging,
8-40 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
πD02
πD 2
1 + fL / D + K L 1 + fL / D + K L − 12 2 gz dt = − dz → dt = − dz = − z dz The last relation 4 1 + fL / D + K L 4 2 gz 2g D D can be integrated easily since the variables are separated. Letting tf be the discharge time and integrating it from t = 0 when z = z1 to t = tf when z = 0 (completely drained tank) gives
∫
tf
dt = −
t =0
1 + fL / D + K L 2g
D02 D
2
∫
D02 2
0
z
−1 / 2
z = z1
dz → t f = -
D02 2
D02 D
2
1
1 + fL / D + K L z 2 1 2g 2
0
= z1
2 D02 D
2
1 + fL / D + K L 12 z1 Simplifying 2g
and substituting the values given, the draining time is determined to be tf =
D02 D
2
2 z1 (1 + fL / D + K L ) (3 m) 2 = g (0.1 m) 2
2(2 m)[1 + (0.015)(100 m)/(0.1 m) + 0.5] 9.81 m/s 2
= 2334 s = 38.9 min
Discussion It can be shown by setting L = 0 that the draining time without the pipe is only 11.7 min. Therefore, the pipe in this case increases the draining time by more than 3 folds.
8-41 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-72 Solution A water tank open to the atmosphere is initially filled with water. A sharp-edged orifice at the bottom drains to the atmosphere through a long pipe equipped with a pump. For a specified initial velocity, the required useful pumping power and the time required to empty the tank are to be determined. Assumptions 1 The flow is uniform and incompressible. 2 The draining pipe is horizontal. 3 The flow is turbulent so that the tabulated value of the loss coefficient can be used. 4 The friction factor remains constant. 5 The effect of the kinetic energy correction factor is negligible, so we set α = 1. Properties The loss coefficient is KL = 0.5 for a sharp-edged entrance. The friction factor of the pipe is given to be 0.015. The density of water at 30°C is ρ = 996 kg/m3. Analysis (a) We take point 1 at the free surface of the tank, and point 2 at the exit of the pipe. We take the reference level at the centerline of the orifice (z2 = 0), and take the positive direction of z to be upwards. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 + hpump, u = α 2
V 22 + hL 2g
where α2 = 1 and ⎛ L + h L = h L,total = h L,major + h L,minor = ⎜ f ⎝ D
∑
2 2 ⎞V ⎛ L ⎞V =⎜ f + KL ⎟ KL ⎟ ⎠ 2g ⎝ D ⎠ 2g
since the diameter of the piping system is constant. Substituting and noting that the initial discharge velocity is 4 m/s, the required useful pumping head and power are determined to be m = ρAcV 2 = ρ (πD 2 / 4 )V 2 = (996 kg/m 3 )[π (0.1 m) 2 /4](4 m/s ) = 31.3 kg/m 3 2 2 L 100 m ⎛ ⎞V ⎛ ⎞ (4 m/s) h pump, u = ⎜1 + f + K L ⎟ 2 − z1 = ⎜1 + (0.015) + 0.5 ⎟ − (2 m) = 11.46 m D 0.1 m ⎝ ⎠ 2g ⎝ ⎠ 2(9.81 m/s 2 )
⎛ ⎞⎛ 1 kW ⎞ 1 kN ⎟ W pump, u = VΔP = m ghpump, u = (31.3 kg/s)(9.81 m/s 2 )(11.46 m)⎜ = 3.52 kW ⎜ 1000 kg ⋅ m/s 2 ⎟⎜⎝ 1 kN ⋅ m/s ⎟⎠ ⎝ ⎠ Therefore, the pump must supply 3.52 kW of mechanical energy to water. Note that the shaft power of the pump must be greater than this to account for the pump inefficiency.
(b) When the discharge velocity remains constant, the flow rate of water becomes V = AcV 2 = (πD 2 / 4 )V 2 = [π (0.1 m) 2 /4](4 m/s ) = 0.03142 m 3 /s
1
The volume of water in the tank is V = Atank z1 = (πD02 / 4) z1 = [π (3 m) 2 /4](2 m) = 14.14 m 3
Then the discharge time becomes V 14.14 m 3 Δt = = = 450 s = 7.5 min V 0.03142 m 3 /s
Water tank 10 m
2m 10 cm 100 m Pump
2
4 m/s
Discussion
1 Note that the pump reduces the discharging time from 38.9 min to 7.5 min. The assumption of constant discharge velocity can be justified on the basis of the pump head being much larger than the elevation head (therefore, the pump will dominate the discharging process). The answer obtained assumes that the elevation head remains constant at 2 m (rather than decreasing to zero eventually), and thus it under predicts the actual discharge time. By an exact analysis, it can be shown that when the effect of the decrease in elevation is considered, the discharge time becomes 468 s = 7.8 min. This is demonstrated below. 2 The required pump head (of water) is 11.46 m, which is more than 10.3 m of water column which corresponds to the atmospheric pressure at sea level. If the pump exit is at 1 atm, then the absolute pressure at pump inlet must be negative ( = -1.16 m or – 11.4 kPa), which is impossible. Therefore, the system cannot work if the pump is installed near the pipe exit,
8-42 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes and cavitation will occur long before the pipe exit where the pressure drops to 4.2 kPa and thus the pump must be installed close to the pipe entrance. A detailed analysis is given below. Demonstration 1 for Prob. 8-72 (extra) (the effect of drop in water level on discharge time) Noting that the water height z in the tank is variable, the average discharge velocity through the pipe at any given time, in general, can be expressed as 2 2 g ( z + hpump, u ) L ⎛ ⎞V + K L ⎟ 2 − z → V2 = hpump, u = ⎜1 + f 1 + fL / D + K L D ⎝ ⎠ 2g where z is the water height relative to the center of the orifice at that time. We denote the diameter of the pipe by D, and the diameter of the tank by D0. The flow rate of water from the tank can be obtained by multiplying the discharge velocity by the cross-sectional area of the pipe,
πD V = A pipeV 2 = 4
2
2 g ( z + hpump, u ) 1 + fL / D + K L
Then the amount of water that flows through the orifice during a differential time interval dt is πD 2 2 g ( z + h pump, u ) dV = Vdt = dt (1) 4 1 + fL / D + K L which, from conservation of mass, must be equal to the decrease in the volume of water in the tank, πD 2 (2) dV = Atan k (−dz ) = − 0 dz 4 where dz is the change in the water level in the tank during dt. (Note that dz is a negative quantity since the positive direction of z is upwards. Therefore, we used –dz to get a positive quantity for the amount of water discharged). Setting Eqs. (1) and (2) equal to each other and rearranging,
πD 2
2 g ( z + hpump,u )
4
1 + fL / D + K L
dt = −
πD02 4
dz
→ dt = −
D02 D2
1 + fL / D + K L −1 ( z + hpump,u ) 2 dz 2g
The last relation can be integrated easily since the variables are separated. Letting tf be the discharge time and integrating it from t = 0 when z = z1 to t = tf when z = 0 (completely drained tank) gives
∫
tf
dt = −
D02 2
1 + fL / D + K L 2g
D Performing the integration gives t =0
1
D 2 1 + fL / D + K L ( z + hpump ) 2 t f = - 02 1 2g D 2
∫
0
= z1
0 z = z1
( z + hpump, u ) −1 / 2 dz
2hpump (1 + fL / D + K L ) ⎞ D02 ⎛⎜ 2( z1 + hpump )(1 + fL / D + K L ) ⎟ Substituting the − 2⎜ ⎟ g g D ⎝ ⎠
values given, the draining time is determined to be (3 m) 2 ⎛ 2(2 + 11.46 m)[1 + 0.015 × 100/0.1 + 0.5] 2(11.46 m)[1 + 0.015 × 100/0.1 + 0.5] ⎞ ⎜ ⎟ − 2 ⎜ 2 ⎟ (0.1 m) ⎝ 9.81 m/s 9.81 m/s 2 ⎠ = 468 s = 7.8 min
tf =
Demonstration 2 for Prob. 8-72 (on cavitation) We take the pump as the control volume, with point 1 at the inlet and point 2 at the exit. We assume the pump inlet and outlet diameters to be the same and the elevation difference between the pump inlet and the exit to be negligible. Then we have z1 ≅ z2 and V1 ≅ V2. The pump is located near the pipe exit, and thus the pump exit pressure is equal to the pressure at the pipe exit, which is the atmospheric pressure, P2 = Patm. Also, the can take hL = 0 since the frictional effects and loses in the pump are accounted for by the pump efficiency. Then the energy equation for the pump (in terms of heads) reduces to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L 2g 2g ρg ρg Solving for P1 and substituting,
→
P1,abs
ρg
+ hpump, u =
Patm ρg
8-43 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes P1,abs = Patm − ρghpump, u ⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎟ = (101.3 kPa) − (996 kg/m 3 )(9.81 m/s 2 )(11.46 m)⎜ = - 10.7 kPa ⎜ 1000 kg ⋅ m/s 2 ⎟⎜⎝ 1 kN/m 2 ⎟⎠ ⎝ ⎠ which is impossible (absolute pressure cannot be negative). The technical answer to the question is that cavitation will occur since the pressure drops below the vapor pressure of 4.246 kPa. The practical answer is that the question is invalid (void) since the system will not work anyway. Therefore, we conclude that the pump must be located near the beginning, not the end of the pipe. Note that when doing a cavitation analysis, we must work with the absolute pressures. (If the system were installed as indicated, a water velocity of V = 4 m/s could not be established regardless of how much pump power were applied. This is because the atmospheric air and water elevation heads alone are not sufficient to drive such flow, with the pump restoring pressure after the flow.) To determine the furthest distance from the tank the pump can be located without allowing cavitation, we assume the pump is located at a distance L* from the exit, and choose the pump and the discharge portion of the pipe (from the pump to the exit) as the system, and write the energy equation. The energy equation this time will be as above, except that hL (the pipe losses) must be considered and the pressure at 1 (pipe inlet) is the cavitation pressure, P1 = 4.246 kPa: P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbine, e + h L 2g 2g ρg ρg
→
P1,abs
ρg
+ hpump, u =
Patm L* V 2 +f ρg D 2g
P1,abs − Patm L* V = + h pump, u D 2g ρg Substituting the given values and solving for L* gives 2
or
f
⎞ ⎟ + (11.46 m) → L* = 12.5 m ⎟ ⎠ Therefore, the pump must be at least 12.5 m from the pipe exit to avoid cavitation at the pump inlet (this is where the lowest pressure occurs in the piping system, and where the cavitation is most likely to occur). Cavitation onset places an upper limit to the length of the pipe on the suction side. A pipe slightly longer would become vapor bound, and the pump could not pull the suction necessary to sustain the flow. Even if the pipe on the suction side were slightly shorter than 100 – 12.5 = 87.5 m, cavitation can still occur in the pump since the liquid in the pump is usually accelerated at the expense of pressure, and cavitation in the pump could erode and destroy the pump. Also, over time, scale and other buildup inside the pipe can and will increase the pipe roughness, increasing the friction factor f, and therefore the losses. Buildup also decreases the pipe diameter, which increases pressure drop. Therefore, flow conditions and system performance may change (generally decrease) as the system ages. A new system that marginally misses cavitation may degrade to where cavitation becomes a problem. Proper design avoids these problems, or where cavitation cannot be avoided for some reason, it can at least be anticipated. (0.015)
(4 m/s) 2 (4.246 − 101.3) kN/m 2 ⎛⎜ 1000 kg ⋅ m/s 2 L* = 0.1 m 2(9.81 m/s 2 ) (996 kg/m 3 )(9.81 m/s 2 ) ⎜⎝ 1 kN
8-44 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-73 Solution Oil is flowing through a vertical glass funnel which is always maintained full. The flow rate of oil through the funnel and the funnel effectiveness are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed (to be verified). 3 The frictional loses in the cylindrical reservoir are negligible since its diameter is very large and thus the oil velocity is very low. Properties
The density and viscosity of oil at 20°C are ρ = 888.1 kg/m3 and μ = 0.8374 kg/m⋅s.
Analysis We take point 1 at the free surface of the oil in the cylindrical reservoir, and point 2 at the exit of the funnel pipe which is also taken as the reference level (z2 = 0). The fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is negligible (V1 ≅ 0). For the ideal case of “frictionless flow,” the exit velocity is determined from the Bernoulli equation to be P1 V12 P V2 + + z1 = 2 + 2 + z 2 ρg 2 g ρg 2 g Substituting,
→
V 2 = V 2,max = 2gz 1
1
V 2,max = 2 gz1 = 2(9.81 m/s 2 )(0.40 m ) = 2.801 m/s
15 cm
Oil
This is the flow velocity for the frictionless case, and thus it is the maximum flow velocity. Then the maximum flow rate and the Reynolds number become Vmax = V 2,max A2 = V 2,max (πD 22 / 4) 1 cm
= (2.801 m/s)[π (0.01 m) 2 / 4] = 2.20 × 10 − 4 m 3 /s
ρVD (888.1 kg/m 3 )(2.801 m/s)(0.01 m) Re = = = 29.71 0.8374 kg/m ⋅ s μ
25 cm
2
which is less than 2300. Therefore, the flow is laminar, as postulated. (Note that in the actual case the velocity and thus the Reynolds number will be even smaller, verifying the flow is always laminar). The entry length in this case is Lh = 0.05 Re D = 0.05 × 29.71× (0.01 m ) = 0.015 m
which is much less than the 0.25 m pipe length. Therefore, the entrance effects can be neglected as postulated. Noting that the flow through the pipe is laminar and can be assumed to be fully developed, the flow rate can be determined from the appropriate relation with θ = -90° since the flow is downwards in the vertical direction, ( ΔP − ρgL sin θ )πD 4 V = 128μL where ΔP = Ppipe inlet − Ppipe exit = ( Patm + ρghcylinder ) − Patm = ρghcylinder is the pressure difference across the pipe, L = hpipe, and sin θ = sin (-90°) = -1. Substituting, the flow rate is determined to be
ρg (hcylinder + hpipe )πD 4
(888.1 kg/m 3 )(9.81 m/s 2 )(0.15 + 0.25 m)π (0.01 m) 4 = 4.09 × 10 -6 m 3 /s 128μL 128(0.8374 kg/m ⋅ s)(0.25 m) Then the “funnel effectiveness” becomes 4.09 × 10 −6 m 3 /s V Eff = = = 0.0186 or 1.86% Vmax 2.20 × 10 − 4 m 3 /s V =
=
Note that the flow is driven by gravity alone, and the actual flow rate is a small fraction of the flow rate that Discussion would have occurred if the flow were frictionless.
8-45 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-74 Solution Oil is flowing through a vertical glass funnel which is always maintained full. The flow rate of oil through the funnel and the funnel effectiveness are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed (to be verified). 3 The frictional loses in the cylindrical reservoir are negligible since its diameter is very large and thus the oil velocity is very low.
The density and viscosity of oil at 20°C are ρ = 888.1 kg/m3 and μ = 0.8374 kg/m⋅s.
Properties
Analysis We take point 1 at the free surface of the oil in the cylindrical reservoir, and point 2 at the exit of the funnel pipe, which is also taken as the reference level (z2 = 0). The fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is negligible (V1 ≅ 0). For the ideal case of “frictionless flow,” the exit velocity is determined from the Bernoulli equation to be P1 V12 P V2 + + z1 = 2 + 2 + z 2 ρg 2 g ρg 2 g
→
V 2 = V 2,max = 2gz 1
(a) Case 1: Pipe length remains constant at 25 cm, but the pipe diameter is doubled to D2 = 2 cm:
1 15 cm
Oil
Substitution gives V 2,max = 2 gz1 = 2(9.81 m/s 2 )(0.40 m ) = 2.801 m/s
This is the flow velocity for the frictionless case, and thus it is the maximum flow velocity. Then the maximum flow rate and the Reynolds number become Vmax = V 2,max A2 = V 2,max (πD 22 / 4) = (2.801 m/s)[π (0.02 m) 2 / 4]
= 8.80 × 10 Re =
−4
2 cm
25 cm
2
3
m /s
ρVD (888.1 kg/m 3 )(2.801 m/s)(0.02 m) = = 59.41 0.8374 kg/m ⋅ s μ
which is less than 2300. Therefore, the flow is laminar. (Note that in the actual case the velocity and thus the Reynolds number will be even smaller, verifying the flow is always laminar). The entry length is Lh = 0.05 Re D = 0.05 × 59.41× (0.02 m ) = 0.059 m
which is considerably less than the 0.25 m pipe length. Therefore, the entrance effects can be neglected (with reservation). Noting that the flow through the pipe is laminar and can be assumed to be fully developed, the flow rate can be determined from the appropriate relation with θ = -90° since the flow is downwards in the vertical direction, (ΔP − ρgL sin θ )πD 4 V = 128μL where ΔP = Ppipe inlet − Ppipe exit = ( Patm + ρghcylinder ) − Patm = ρghcylinder is the pressure difference across the pipe, L = hpipe, and sin θ = sin (-90°) = -1. Substituting, the flow rate is determined to be V =
ρg (hcylinder + hpipe )πD 4 128μL
=
(888.1 kg/m 3 )(9.81 m/s 2 )(0.15 + 0.25 m)π (0.02 m) 4 = 6.54 × 10 - 5 m 3 /s 128(0.8374 kg/m ⋅ s)(0.25 m)
Then the “funnel effectiveness” becomes Eff =
V V
max
=
0.654 × 10 −4 m 3 /s 8.80 × 10 − 4 m 3 /s
= 0.074
or
7.4%
(b) Case 2: Pipe diameter remains constant at 1 cm, but the pipe length is doubled to L = 50 cm:
15 cm
Oil
Substitution gives V 2,max = 2 gz1 = 2(9.81 m/s 2 )(0.65 m ) = 3.571 m/s
1 cm
50 cm
8-46 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes This is the flow velocity for the frictionless case, and thus it is the maximum flow velocity. Then the maximum flow rate and the Reynolds number become Vmax = V 2,max A2 = V 2,max (πD 22 / 4) = (3.571 m/s)[π (0.01 m) 2 / 4]
= 2.805 × 10 − 4 m 3 / s
Re =
ρVD (888.1 kg/m 3 )(3.571 m/s)(0.01 m) = = 37.87 0.8374 kg/m ⋅ s μ
which is less than 2300. Therefore, the flow is laminar. (Note that in the actual case the velocity and thus the Reynolds number will be even smaller, verifying the flow is always laminar). The entry length is L h = 0.05 Re D = 0.05 × 37.87 × (0.01 m ) = 0.019 m
which is much less than the 0.50 m pipe length. Therefore, the entrance effects can be neglected. Noting that the flow through the pipe is laminar and can be assumed to be fully developed, the flow rate can be determined from the appropriate relation with θ = -90° since the flow is downwards in the vertical direction, (ΔP − ρgL sin θ )πD 4 V = 128μL where ΔP = Ppipe inlet − Ppipe exit = ( Patm + ρghcylinder ) − Patm = ρghcylinder is the pressure difference across the pipe, L = hpipe, and sin θ = sin (-90°) = -1. Substituting, the flow rate is determined to be V =
ρg (hcylinder + hpipe )πD 4 128μL
=
(888.1 kg/m 3 )(9.81 m/s 2 )(0.15 + 0.50 m)π (0.01 m) 4 = 3.32 × 10 - 6 m 3 /s 128(0.8374 kg/m ⋅ s)(0.50 m)
Then the “funnel effectiveness” becomes Eff =
V V
max
=
3.32 × 10 −6 m 3 /s 2.805 × 10 − 4 m 3 /s
= 0.0118
or
1.18%
Discussion Note that the funnel effectiveness increases as the pipe diameter is increased, and decreases as the pipe length is increased. This is because the frictional losses are proportional to the length but inversely proportional to the diameter of the flow sections.
8-47 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-75 Solution Water is drained from a large reservoir through two pipes connected in series. The discharge rate of water from the reservoir is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The pipes are horizontal. 3 The entrance effects are negligible, and thus the flow is fully developed. 4 The flow is turbulent so that the tabulated value of the loss coefficients can be used. 5 The pipes involve no components such as bends, valves, and other connectors. 6 The piping section involves no work devices such as pumps and turbines. 7 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 8 The water level in the reservoir remains constant. 9 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 15°C are ρ = 999.1 kg/m3 and μ = 1.138×10-3 kg/m⋅s, respectively. The loss coefficient is KL = 0.5 for a sharp-edged entrance, and it is 0.46 for the sudden contraction, corresponding to d2/D2 = 42/102 = 0.16. The pipes are made of plastic and thus they are smooth, ε = 0. Analysis We take point 1 at the free surface of the reservoir, and point 2 at the exit of the pipe, which is also taken to be the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm), the fluid level in the reservoir is constant (V1 = 0), and that there are no work devices such as pumps and turbines, the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g
where α2 = 1. Substituting, V 22 18 m = + hL 2(9.81m/s 2 ) where
→
z1 = α 2
V 22 + hL 2g
1 (1) 18 m
Water tank
h L = h L, total = h L ,major + h L,minor =
⎛ L + ⎜f ⎝ D
∑
∑K
2
L
2
⎞V ⎟ ⎠ 2g 35 m
20 m
Note that the diameters of the two pipes, and thus the flow velocities through them are different. Denoting the first pipe by 1 and the second pipe by 2, and using conservation of mass, the velocity in the first pipe can be expressed in terms of V2 as → ρV1 A1 = ρV 2 A2 → V1 =
m 1 = m 2
A2 D2 (4 cm ) 2 V 2 = 22 V 2 = V 2 → V1 = 0.16V 2 A1 D1 (10 cm ) 2
Then the head loss can be expressed as ⎛ L ⎞V 2 ⎛ ⎞V 2 L h L = ⎜⎜ f 1 1 + K L ,entrance ⎟⎟ 1 + ⎜⎜ f 2 2 + K L ,contraction ⎟⎟ 2 D2 ⎝ D1 ⎠ 2g ⎝ ⎠ 2g or V12 V 22 20 m 35 m ⎛ ⎞ ⎛ ⎞ hL = ⎜ f 1 f + 0.5 ⎟ + + 0 . 46 ⎜ ⎟ 2 ⎝ 0.10 m ⎠ 2(9.81m/s 2 ) ⎝ 0.04 m ⎠ 2(9.81m/s 2 )
(2)
(3)
The flow rate, the Reynolds number, and the friction factor are expressed as V = V 2 A2 = V 2 (πD 22 / 4)
Re 1 =
ρV1 D1 μ
Re 2 =
ρV 2 D 2 μ
→ Re 1 =
(999.1 kg/m 3 )V1 (0.10 m)
→ Re 2 =
1.138 × 10 − 3 kg/m ⋅ s
(999.1 kg/m 3 )V2 (0.04 m) 1.138 × 10 − 3 kg/m ⋅ s
⎛ε / D 2.51 1 = −2.0 log⎜ + ⎜ 3 . 7 f1 Re 1 f 1 ⎝
1
1 f2
V = V 2 [π (0.04 m) 2 / 4]
→
⎞ ⎟ → ⎟ ⎠
⎛ε / D 2.51 2 = −2.0 log⎜ + ⎜ 3.7 Re 2 f 2 ⎝
(4)
(5) (6)
⎛ 2.51 = −2.0 log⎜ 0 + ⎜ f1 Re1 f 1 ⎝
⎞ ⎟ ⎟ ⎠
(7)
⎞ ⎛ ⎟ → 1 = −2.0 log⎜ 0 + 2.51 ⎟ ⎜ f2 Re 2 f 2 ⎠ ⎝
⎞ ⎟ ⎟ ⎠
(8)
1
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Chapter 8 Flow in Pipes This is a system of 8 equations in 8 unknowns, and their simultaneous solution by an equation solver gives V = 0.00595 m3 /s , V1 = 0.757 m/s, V2 = 4.73 m/s, hL = hL1 + hL2 = 0.13 + 16.73 = 16.86 m,
Re1 = 66,500,
Re2 = 166,200, f1 = 0.0196, f2 = 0.0162
Note that Re > 4000 for both pipes, and thus the assumption of turbulent flow is valid. Discussion This problem can also be solved by using an iterative approach by assuming an exit velocity, but it will be very time consuming. Equation solvers such as EES are invaluable for this kind of problems.
8-49 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-76E Solution The flow rate through a piping system between a river and a storage tank is given. The power input to the pump is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 The elevation difference between the free surfaces of the tank and the river remains constant. 5 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The roughness of galvanized iron pipe is ε = 0.0005 ft. Analysis The piping system involves 125 ft of 5-in diameter piping, an entrance with negligible loses, 3 standard flanged 90° smooth elbows (KL = 0.3 each), and a sharp-edged exit (KL = 1.0). We choose points 1 and 2 at the free surfaces of the river and the tank, respectively. We note that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm), and the fluid velocity is 6 ft/s at point 1 and zero at point 2 (V1 = 6 ft/s and V2 =0). We take the free surface of the river as the reference level (z1 = 0). Then the energy equation for a control volume between these two points simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g
α1
→
V12 + hpump, u = z 2 + h L 2g
where α1 = 1 and
2
2 ⎛ L ⎞V h L = h L, total = h L ,major + h L ,minor = ⎜ f KL ⎟ 2 + ⎝ D ⎠ 2g since the diameter of the piping system is constant. The average velocity in the pipe and the Reynolds number are
∑
V=
125 ft 12 ft
1.5 ft 3 /s V V = = = 11.0 ft/s Ac πD 2 / 4 π (5 / 12 ft) 2 / 4
Re =
5 in
1
ρVD (62.3 lbm/ft )(11.0 ft/s)(5/12 ft) = = 435,500 μ 6.556 × 10 − 4 lbm/ft ⋅ s
Water tank
1.5 ft3/s
3
River
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 0.0005 ft ε/D= = 0.0012 5 / 12 ft The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 0.0012 ⎞ 1 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 Re f ⎟ ⎜ 3. 7 ⎟ f f 435 , 500 f ⎝ ⎠ ⎝ ⎠ It gives f = 0.0211. The sum of the loss coefficients is
∑K
L
= K L,entrance + 3K L,elbow + K L,exit = 0 + 3 × 0.3 + 1.0 = 1.9
Then the total head loss becomes 2 2 125 ft ⎛ L ⎞V ⎛ ⎞ (11.0 ft/s) + = ⎜ (0.0211) + 1.90 ⎟ = 15.5 ft hL = ⎜ f KL ⎟ 5/12 ft ⎝ D ⎠ 2g ⎝ ⎠ 2(32.2 ft/s 2 ) The useful pump head input and the required power input to the pump are
∑
h pump, u = z 2 + h L −
Wpump =
Wpump, u
ηpump
=
V12 (6 ft/s) 2 = 12 + 15.5 − = 26.9 ft 2g 2(32.2 ft/s 2 )
V ρ ghpump, u
ηpump
(1.5 ft /s )( 62.30 lbm/ft )( 32.2 ft/s ) ( 26.9 ft ) ⎛ 3
3
2
1 lbf ⎞⎛ 1 kW ⎞ ⎜ ⎟ = 4.87 kW 2 ⎟⎜ 0.70 ⎝ 32.2 lbm ⋅ ft/s ⎠⎝ 737 lbf ⋅ ft/s ⎠ Therefore, 4.87 kW of electric power must be supplied to the pump. =
8-50 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.0211, which is identical to the calculated value. The friction coefficient would drop to 0.0135 if smooth pipes were used. Note that fL/D = 6.3 in this case, which is about 3 times the total minor loss coefficient of 1.9. Therefore, the frictional losses in the pipe dominate the minor losses, but the minor losses are still significant.
8-77 Solution In the previous problem, the effect of the pipe diameter on pumping power for the same constant flow rate is to be investigated by varying the pipe diameter from 1 in to 10 in in increments of 1 in. The EES Equations window is printed below, along with the tabulated and plotted results.
g=32.2 L= 125 D=Dinch/12 z2= 12 rho=62.30 nu=mu/rho mu=0.0006556 eff=0.70 Re=V2*D/nu A=pi*(D^2)/4 V2=Vdot/A Vdot= 1.5 V1=6 eps1=0.0005 rf1=eps1/D 1/sqrt(f1)=-2*log10(rf1/3.7+2.51/(Re*sqrt(f1))) KL= 1.9 HL=(f1*(L/D)+KL)*(V2^2/(2*g)) hpump=z2+HL-V1^2/(2*32.2) Wpump=(Vdot*rho*hpump)/eff/737
12000 10000 8000
Wpump, kW
Analysis
6000 4000 2000 0 1
2
3
4
5
6
7
8
9
10
D, in
D, in
Wpump, kW
V, ft/s
Re
1 2 3 4 5 6 7 8 9 10
2.178E+06 1.089E+06 7.260E+05 5.445E+05 4.356E+05 3.630E+05 3.111E+05 2.722E+05 2.420E+05 2.178E+05
275.02 68.75 30.56 17.19 11.00 7.64 5.61 4.30 3.40 2.75
10667.48 289.54 38.15 10.55 4.88 3.22 2.62 2.36 2.24 2.17
Discussion We see that the required pump power decreases very rapidly as pipe diameter increases. This is due to the significant decrease in irreversible head loss in larger diameter pipes.
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Chapter 8 Flow in Pipes 8-78 Solution A solar heated water tank is to be used for showers using gravity driven flow. For a specified flow rate, the elevation of the water level in the tank relative to showerhead is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 The elevation difference between the free surface of water in the tank and the shower head remains constant. 5 There are no pumps or turbines in the piping system. 6 The losses at the entrance and at the showerhead are said to be negligible. 7 The water tank is open to the atmosphere. 8 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 40°C are ρ = 992.1 kg/m3 and μ = 0.653×10-3 kg/m⋅s, respectively. The loss coefficient is KL = 0.5 for a sharp-edged entrance. The roughness of galvanized iron pipe is ε = 0.00015 m. Analysis The piping system involves 20 m of 1.5-cm diameter piping, an entrance with negligible loss, 4 miter bends (90°) without vanes (KL = 1.1 each), and a wide open globe valve (KL = 10). We choose point 1 at the free surface of water in the tank, and point 2 at the shower exit, which is also taken to be the reference level (z2 = 0). The fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm), and V1 = 0. Then the energy equation for a control volume between these two points simplifies to P1 V2 P V2 V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L → z1 = α 2 2 + h L ρg ρg 2g 2g 2g 2
⎛ L ⎞V KL ⎟ 2 + where h L = h L, total = h L ,major + h L ,minor = ⎜ f ⎝ D ⎠ 2g since the diameter of the piping system is constant. The average velocity in the pipe and the Reynolds number are 0.0007 m 3 /s V V = = = 3.961 m/s V2 = 2 Ac πD / 4 π (0.015 m) 2 / 4
∑
ρV 2 D (992.1 kg/m 3 )(3.961 m/s)(0.015 m) Re = = = 90,270 μ 0.653 × 10 − 3 kg/m ⋅ s
1
Water tank
z1
1.5 cm 20 m
0.7 L/s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 0.00015 m Showers 2 ε/D= = 0.01 0.015 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ 0.0051 ⎛ε / D 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ → = −2.0 log ⎜ + = −2.0 log⎜ + ⎜ ⎜ ⎟ f f 90,270 f ⎟⎠ ⎝ 3. 7 ⎝ 3.7 Re f ⎠ It gives f = 0.03857. The sum of the loss coefficients is K L = K L,entrance + 4 K L,elbow + K L, valve + K L,exit = 0 + 4 × 1.1 + 10 + 0 = 14.4
∑
Note that we do not consider the exit loss unless the exit velocity is dissipated within the system considered (in this case it is not). Then the total head loss and the elevation of the source become 2 2 20 m ⎛ L ⎞V ⎛ ⎞ (3.961 m/s) + + 14.4 ⎟ = 52.6 m hL = ⎜ f K L ⎟ 2 = ⎜ (0.03857) 0.015 m ⎝ D ⎠ 2g ⎝ ⎠ 2(9.81 m/s 2 )
∑
z1 = α 2
V 22 (3.961 m/s) 2 + h L = (1) + 52.6 m = 53.4 m 2g 2(9.81 m/s 2 )
since α2 = 1. Therefore, the free surface of the tank must be 53.4 m above the shower exit to ensure water flow at the specified rate. Discussion
We neglected the minor loss associated with the shower head. In reality, this loss is most likely significant.
8-52 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-79 Solution The flow rate through a piping system connecting two water reservoirs with the same water level is given. The absolute pressure in the pressurized reservoir is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 There are no pumps or turbines in the piping system. Properties The density and dynamic viscosity of water at 10°C are ρ = 999.7 kg/m3 and μ = 1.307×10-3 kg/m⋅s. The loss coefficient is KL = 0.5 for a sharp-edged entrance, KL = 2 for swing check valve, KL = 0.2 for the fully open gate valve, and KL = 1 for the exit. The roughness of cast iron pipe is ε = 0.00026 m. Analysis We choose points 1 and 2 at the free surfaces of the two reservoirs. We note that the fluid velocities at both points are zero (V1 = V2 =0), the fluid at point 2 is open to the atmosphere (and thus P2 = Patm), both points are at the same level (z1 = z2). Then the energy equation for a control volume between these two points simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L → ρg ρg 2g 2g 2 L ⎞V where h L = h L, total = h L ,major + h L ,minor = ⎛⎜ f KL ⎟ 2 + ⎝ D ⎠ 2g since the diameter of the piping system is constant. The average flow velocity and the Reynolds number are 0.0012 m 3 /s V V = = = 3.82 m/s V2 = 2 Ac πD / 4 π (0.02 m) 2 / 4
P1 Patm = + hL ρg ρg
→ P1 = Patm + ρgh L
∑
Re =
ρV 2 D (999.7 kg/m 3 )(3.82 m/s)(0.02 m) = = 58,400 μ 1.307 × 10 − 3 kg/m ⋅ s
Air
P
1
Water
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 0.00026 m = 0.013 ε/D= 0.02 m The friction factor can be determined from the Moody chart, but to avoid the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 0.013 1 2.51 ⎞⎟ 1 2.51 = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ ⎟ ⎜ 3 . 7 3 . 7 f Re f ⎠ f 58,400 f ⎝ ⎝ It gives f = 0.0424. The sum of the loss coefficients is
∑K
L
2
Water
40 m 2 cm
1.2 L/s
reading error, we determine it from the ⎞ ⎟ ⎟ ⎠
= K L,entrance + K L ,check valve + K L, gate valve + K L,exit = 0.5 + 2 + 0.2 + 1 = 3.7
Then the total head loss becomes 2 2 40 m ⎛ L ⎞V ⎛ ⎞ (3.82 m/s) + + 3.7 ⎟ = 65.8 m hL = ⎜ f K L ⎟ 2 = ⎜ (0.0424 ) 0.02 m ⎝ D ⎠ 2g ⎝ ⎠ 2(9.81 m/s 2 ) Substituting, ⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎟ P1 = Patm + ρgh L = (88 kPa) + (999.7 kg/m 3 )(9.81 m/s 2 )(65.8 m)⎜ = 734 kPa ⎜ 1000 kg ⋅ m/s 2 ⎟⎜⎝ 1 kN/m 2 ⎟⎠ ⎝ ⎠ Discussion The absolute pressure above the first reservoir must be 734 kPa, which is quite high. Note that the minor losses in this case are negligible (about 4% of total losses). Also, the friction factor could be determined easily from the explicit Haaland relation (it gives the same result, 0.0424). The friction coefficient would drop to 0.0202 if smooth pipes were used.
∑
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Chapter 8 Flow in Pipes 8-80 Solution A tanker is to be filled with fuel oil from an underground reservoir using a plastic hose. The required power input to the pump is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 Fuel oil level remains constant. 5 Reservoir is open to the atmosphere. Properties The density and dynamic viscosity of fuel oil are given to be ρ = 920 kg/m3 and μ = 0.045 kg/m⋅s. The loss coefficient is KL = 0.12 for a slightly-rounded entrance and KL = 0.3 for a 90° smooth bend (flanged). The plastic pipe is smooth and thus ε = 0. The kinetic energy correction factor at hose discharge is given to be α = 1.05. Analysis We choose point 1 at the free surface of oil in the reservoir and point 2 at the exit of the hose in the tanker. We note the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and the fluid velocity at point 1 is zero (V1 = 0). We take the free surface of the reservoir as the reference level (z1 = 0). Then the energy equation for a control volume between these two points simplifies to P1 V2 P V2 V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L → hpump, u = α 2 2 + z 2 + h L ρg ρg 2g 2g 2g where 2 ⎛ L ⎞V h L = h L, total = h L ,major + h L ,minor = ⎜ f KL ⎟ 2 + ⎝ D ⎠ 2g since the diameter of the piping system is constant. The flow rate is determined from the requirement that the tanker must be filled in 30 min, 2 V 18 m 3 = 0.01 m 3 /s V = tanker = Tanker Δt (30 × 60 s) 18 m3 5 cm Then the average velocity in the pipe and the Reynolds 5m number become 0.01 m 3 /s V V 1 20 m = = = 5.093 m/s V2 = Ac πD 2 / 4 π (0.05 m) 2 / 4
∑
ρV 2 D (920 kg/m 3 )(5.093 m/s)(0.05 m) Re = = = 5206 0.045 kg/m ⋅ s μ
Fuel oil
Pump
which is greater than 4000. Therefore, the flow is turbulent. The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation, ⎛ε / D ⎛ 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ 0 + ⎜ 3.7 Re f ⎟ ⎜ f f 5206 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.0370. The sum of the loss coefficients is K L = K L,entrance + 2 K L,bend = 0.12 + 2 × 0.3 = 0.72
∑
Note that we do not consider the exit loss unless the exit velocity is dissipated within the system (in this case it is not). Then the total head loss, the useful pump head, and the required pumping power become 2 2 20 m ⎛ L ⎞V ⎛ ⎞ (5.093 m/s) + + 0.72 ⎟ = 20.5 m hL = ⎜ f K L ⎟ 2 = ⎜ (0.0370) 0.05 m ⎝ D ⎠ 2g ⎝ ⎠ 2(9.81 m/s 2 )
∑
hpump, u =
W pump =
V 22 (5.093 m/s) 2 + z 2 + h L = 1.05 + 5 m + 20.5 m = 26.9 m 2g 2(9.81 m/s 2 )
Vρghpump, u
η pump
=
(0.01 m 3 /s)(920 kg/m 3 )(9.81 m/s 2 )(26.9 m ) ⎛ 1 kN ⎜ 2 ⎜ 0.82 ⎝ 1000 kg ⋅ m/s
⎞⎛ 1 kW ⎞ ⎟⎜ ⎟⎝ 1 kN ⋅ m/s ⎟⎠ = 2.96 kW ⎠
Discussion Note that the minor losses in this case are negligible (0.72/15.52 = 0.046 or about 5% of total losses). Also, the friction factor could be determined easily from the Haaland relation (it gives 0.0372).
8-54 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-81 Solution Two pipes of identical length and material are connected in parallel. The diameter of one of the pipes is twice the diameter of the other. The ratio of the flow rates in the two pipes is to be determined Assumptions 1 The flow is steady and incompressible. 2 The friction factor is given to be the same for both pipes. 3 The minor losses are negligible. Analysis When two pipes are parallel in a piping system, the head loss for each pipe must be same. When the minor losses are disregarded, the head loss for fully developed flow in a pipe of length L and diameter D can be expressed as hL = f
2 2 L V2 L 1 ⎛ V ⎞ L 1 ⎛ V ⎞ L 1 V 2 L V 2 ⎜ ⎟ =8f ⎜ ⎟ = f = f = 8 f D 2g D 2 g ⎜⎝ Ac ⎟⎠ D 2 g ⎜⎝ πD 2 / 4 ⎟⎠ D g π 2D4 gπ 2 D 5
Solving for the flow rate gives
π 2 h L g 2.5 V = D = kD 2.5 8 fL
(k = constant of proportionality)
When the pipe length, friction factor, and the head loss is constant, which is the case here for parallel connection, the flow rate becomes proportional to the 2.5th power of diameter. Therefore, when the diameter is doubled, the flow rate will increase by a factor of 22.5 = 5.66 since 2D
If
VA = kD A2.5
Then
VB = kD B2.5 = k (2 D A ) 2.5 = 2 2.5 kD A2.5 = 2 2.5VA = 5.66VA
Therefore, the ratio of the flow rates in the two pipes is 5.66.
B L
Discussion
L
A
D
The relationship of flow rate to pipe diameter is not linear or even quadratic.
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Chapter 8 Flow in Pipes 8-82 Solution Cast iron piping of a water distribution system involves a parallel section with identical diameters but different lengths. The flow rate through one of the pipes is given, and the flow rate through the other pipe is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses are negligible. 4 The flow is fully turbulent and thus the friction factor is independent of the Reynolds number (to be verified). Properties The density and dynamic viscosity of water at 15°C are ρ = 999.1 kg/m3 and μ = 1.138×10-3 kg/m⋅s. The roughness of cast iron pipe is ε = 0.00026 m.
The average velocity in pipe A is 0.4 m3/s 30 cm 1000 m V V 0.4 m 3 /s VA = = = = 5 . 659 m/s A Ac πD 2 / 4 π (0.30 m) 2 / 4 When two pipes are parallel in a piping system, the head loss for each pipe must be same. When the minor losses are disregarded, the head loss for fully developed flow in a pipe of length L and diameter D is B 30 cm 1000 m L V2 hL = f D 2g Writing this for both pipes and setting them equal to each other, and noting that DA = DB (given) and fA = fB (to be verified) gives Analysis
fA
L A V A2 L V2 = fB B B D A 2g DB 2g
→
VB = V A
LA 1000 m = (5.659 m/s) = 3.267 m/s 3000 m LB
Then the flow rate in pipe B becomes VB = AcV B = [πD 2 / 4]V B = [π (0.30 m) 2 / 4](3.267 m/s) = 0.231 m 3 /s
Proof that flow is fully turbulent and thus friction factor is independent of Reynolds number:
The velocity in pipe B is lower. Therefore, if the flow is fully turbulent in pipe B, then it is also fully turbulent in pipe A. The Reynolds number in pipe B is Re B =
ρV B D (999.1 kg/m 3 )(3.267 m/s)(0.30 m) = = 0.860 × 10 6 μ 1.138 × 10 − 3 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is
ε/D=
0.00026 m = 0.00087 0.30 m
From Moody’s chart, we observe that for a relative roughness of 0.00087, the flow is fully turbulent for Reynolds number greater than about 106. Therefore, the flow in both pipes is fully turbulent, and thus the assumption that the friction factor is the same for both pipes is valid. Discussion Note that the flow rate in pipe B is less than the flow rate in pipe A because of the larger losses due to the larger length.
8-56 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-83 Solution Cast iron piping of a water distribution system involves a parallel section with identical diameters but different lengths and different valves. The flow rate through one of the pipes is given, and the flow rate through the other pipe is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses other than those for the valves are negligible. 4 The flow is fully turbulent and thus the friction factor is independent of the Reynolds number. Properties The density and dynamic viscosity of water at 15°C are ρ = 999.1 kg/m3 and μ = 1.138×10-3 kg/m⋅s. The roughness of cast iron pipe is ε = 0.00026 m. Analysis For pipe A, the average velocity and the Reynolds number are V VA 0.4 m 3 /s = = 5.659 m/s VA = A = Ac πD 2 / 4 π (0.30 m) 2 / 4 Re A =
0.4 m3/s
30 cm
1000 m
A
ρV A D (999.1 kg/m 3 )(5.659 m/s)(0.30 m) = = 1.49 × 10 6 −3 μ 1.138 × 10 kg/m ⋅ s
The relative roughness of the pipe is
B 1000 m
0.00026 m ε/D= = 8.667 × 10 − 4 0.30 m The friction factor corresponding to this relative roughness and the Reynolds number can Moody chart. To avoid the reading error, we determine it from the Colebrook equation ⎛ε / D ⎛ 8.667 × 10 −4 1 2.51 ⎞⎟ 1 2.51 = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ ⎟ ⎜ 3 . 7 3 . 7 f Re f ⎠ f 1.49 × 10 6 f ⎝ ⎝
30 cm
simply be determined from the ⎞ ⎟ ⎟ ⎠
It gives f = 0.0192. Then the total head loss in pipe A becomes 2 2 1000 m ⎛ L ⎞V ⎛ ⎞ (5.659 m/s) h L , A = ⎜ f A + K L ⎟ A = ⎜ (0.0192) + 2.1⎟ = 107.9 m 0.30 m ⎠ 2(9.81 m/s 2 ) ⎝ D ⎠ 2g ⎝ When two pipes are parallel in a piping system, the head loss for each pipe must be same. Therefore, the head loss for pipe B must also be 107.9 m. Then the average velocity in pipe B and the flow rate become 2 ⎛ L ⎞V h L, B = ⎜ f B + K L ⎟ B ⎝ D ⎠ 2g
V B2 3000 m ⎛ ⎞ → 107.9 m = ⎜ (0.0192) + 10 ⎟ 0.30 m ⎝ ⎠ 2(9.81 m/s 2 )
→ V B = 3.24 m/s
VB = AcV B = [πD 2 / 4]V B = [π (0.30 m) 2 / 4](3.24 m/s) = 0.229 m 3 /s
Discussion Note that the flow rate in pipe B decreases slightly (from 0.231 to 0.229 m3/s) due to the larger minor loss in that pipe. Also, minor losses constitute just a few percent of the total loss, and they can be neglected if great accuracy is not required.
8-57 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-84 Solution Geothermal water is supplied to a city through stainless steel pipes at a specified rate. The electric power consumption and its daily cost are to be determined, and it is to be assessed if the frictional heating during flow can make up for the temperature drop caused by heat loss. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses are negligible because of the large length-to-diameter ratio and the relatively small number of components that cause minor losses. 4 The geothermal well and the city are at about the same elevation. 5 The properties of geothermal water are the same as fresh water. 6 The fluid pressures at the wellhead and the arrival point in the city are the same. Properties The properties of water at 110°C are ρ = 950.6 kg/m3, μ = 0.255×10-3 kg/m⋅s, and Cp = 4.229 kJ/kg⋅°C. The roughness of stainless steel pipes is 2×10-6 m. Analysis (a) We take point 1 at the well-head of geothermal resource and point 2 at the final point of delivery at the city, and the entire piping system as the control volume. Both points are at the same elevation (z2 = z2) and the same velocity (V1 = V2) since the pipe diameter is constant, and the same pressure (P1 = P2). Then the energy equation for this control volume simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g That is, the pumping power is to be used to overcome the head losses due to friction in flow. The average velocity and the Reynolds number are 1.5 m 3 /s V V = = = 5.305 m/s V= Ac πD 2 / 4 π (0.60 m) 2 / 4 Re =
→
h pump, u = h L
2
1 Water
D = 60 cm
1.5 m3/s
L = 12 km
ρVD (950.6 kg/m 3 )(5.305 m/s)(0.60 m) = = 1.187 × 10 7 μ 0.255 × 10 − 3 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 2 × 10 −6 m ε/D= = 3.33 × 10 −6 0.60 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ 3.33 × 10 −6 ⎞ ⎛ε / D 1 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + 7 ⎜ ⎟ ⎜ 3.7 Re f ⎟ 3.7 f f 1 . 187 10 f × ⎝ ⎠ ⎝ ⎠ It gives f = 0.00829. Then the pressure drop, the head loss, and the required power input become ΔP = ΔPL = f hL =
L ρV 2 12,000 m (950.6 kg/m 3 )(5.305 m/s) 2 = 0.00829 D 2 0.60 m 2
⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎜⎜ ⎟⎟⎜ = 2218 kPa 2 ⎟ ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN/m ⎠
ΔPL L V2 12,000 m (5.305 m/s) 2 = f = (0.00829) = 238 m D 2g 0.60 m 2(9.81 m/s 2 ) ρg Welectric, in =
Wpump, u
η pump-motor
=
V ΔP
η pump-motor
(1.5 m /s ) ( 2218 kPa ) ⎛ = 3
1 kW ⎞ ⎜ ⎟ = 4496 kW ≅ 4500 kW 3 ⎝ 1 kPa ⋅ m /s ⎠
0.74
Therefore, the pumps will consume 4496 kW of electric power to overcome friction and maintain flow. The pumps must raise the pressure of the geothermal water by 2218 kPa. Providing a pressure rise of this magnitude at one location may create excessive stress in piping at that location. Therefore, it is more desirable to raise the pressure by smaller amounts at a several locations along the flow. This will keep the maximum pressure in the system and the stress in piping at a safer level. (b) The daily cost of electric power consumption is determined by multiplying the amount of power used per day by the unit cost of electricity, Amount = W elect,in Δt = (4496 kW)(24 h/day) = 107,900 kWh/day Cost = Amount × Unit cost = (107 ,900 kWh/day )( $0.06/kWh ) = $6474/day ≅ $6470/day
8-58 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes (c) The energy consumed by the pump (except the heat dissipated by the motor to the air) is eventually dissipated as heat due to the frictional effects. Therefore, this problem is equivalent to heating the water by a 4496 kW of resistance heater (again except the heat dissipated by the motor). To be conservative, we consider only the useful mechanical energy supplied to the water by the pump. The temperature rise of water due to this addition of energy is W mech = ρVc p ΔT → ΔT =
η pump-motorW elect,in ρVc p
=
0.74 × (4496 kJ/s) (950.6 kg/m 3 )(1.5 m 3 /s)(4.229 kJ/kg ⋅ °C)
= 0.55°C
Therefore, the temperature of water will rise at least 0.55°C, which is more than the 0.5°C drop in temperature (in reality, the temperature rise will be more since the energy dissipation due to pump inefficiency will also appear as temperature rise of water). Thus we conclude that the frictional heating during flow can more than make up for the temperature drop caused by heat loss. Discussion The pumping power requirement and the associated cost can be reduced by using a larger diameter pipe. But the cost savings should be compared to the increased cost of larger diameter pipe.
8-59 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-85 Solution Geothermal water is supplied to a city through cast iron pipes at a specified rate. The electric power consumption and its daily cost are to be determined, and it is to be assessed if the frictional heating during flow can make up for the temperature drop caused by heat loss. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses are negligible because of the large length-to-diameter ratio and the relatively small number of components that cause minor losses. 4 The geothermal well and the city are at about the same elevation. 5 The properties of geothermal water are the same as fresh water. 6 The fluid pressures at the wellhead and the arrival point in the city are the same. Properties The properties of water at 110°C are ρ = 950.6 kg/m3, μ = 0.255×10-3 kg/m⋅s, and Cp = 4.229 kJ/kg⋅°C. The roughness of cast iron pipes is 0.00026 m. Analysis (a) We take point 1 at the well-head of geothermal resource and point 2 at the final point of delivery at the city, and the entire piping system as the control volume. Both points are at the same elevation (z2 = z2) and the same velocity (V1 = V2) since the pipe diameter is constant, and the same pressure (P1 = P2). Then the energy equation for this control volume simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g That is, the pumping power is to be used to overcome the head losses due to friction in flow. The average velocity and the Reynolds number are 1.5 m 3 /s V V = = = 5.305 m/s V= Ac πD 2 / 4 π (0.60 m) 2 / 4 Re =
ρVD (950.6 kg/m 3 )(5.305 m/s)(0.60 m) = = 1.187 × 10 7 μ 0.255 × 10 − 3 kg/m ⋅ s
→
h pump, u = h L
1
2
Water
D = 60 cm
1.5 m3/s
L = 12 km
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 0.00026 m ε/D= = 4.33 × 10 − 4 0.60 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ 4.33 × 10 −4 ⎞ ⎛ε / D 1 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + 7 ⎜ ⎟ ⎜ 3.7 Re f ⎟ 3.7 f f 1 . 187 10 f × ⎝ ⎠ ⎝ ⎠ It gives f = 0.0162. Then the pressure drop, the head loss, and the required power input become ΔP = ΔPL = f hL =
L ρV 2 12,000 m (950.6 kg/m 3 )(5.305 m/s) 2 = 0.0162 D 2 0.60 m 2
⎛ ⎞⎛ 1 kPa ⎞ 1 kN ⎜⎜ ⎟⎟⎜ ⎟ = 4334 kPa 2 ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN/m ⎠
ΔPL L V2 12,000 m (5.305 m/s) 2 = f = (0.0162) = 465 m D 2g 0.60 m 2(9.81 m/s 2 ) ρg Welect, in =
Wpump, u
η pump-motor
=
V ΔP
η pump-motor
(1.5 m /s ) ( 4334 kPa ) ⎛ 3
=
1 kW ⎞ ⎜ ⎟ = 8785 kW ≅ 8790 kW 3 ⎝ 1 kPa ⋅ m /s ⎠
0.74
Therefore, the pumps will consume 8785 kW of electric power to overcome friction and maintain flow. The pumps must raise the pressure of the geothermal water by 4334 kPa. Providing a pressure rise of this magnitude at one location may create excessive stress in piping at that location. Therefore, it is more desirable to raise the pressure by smaller amounts at a several locations along the flow. This will keep the maximum pressure in the system and the stress in piping at a safer level. (b) The daily cost of electric power consumption is determined by multiplying the amount of power used per day by the unit cost of electricity, Amount = W elect,in Δt = (8785 kW)(24 h/day) = 210,800 kWh/day Cost = Amount × Unit cost = ( 210 ,800 kWh/day )( $0.06/kWh ) = $12,650/day ≅ $12,700/day
(c) The energy consumed by the pump (except the heat dissipated by the motor to the air) is eventually dissipated as heat
8-60 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes due to the frictional effects. Therefore, this problem is equivalent to heating the water by a 8785 kW of resistance heater (again except the heat dissipated by the motor). To be conservative, we consider only the useful mechanical energy supplied to the water by the pump. The temperature rise of water due to this addition of energy is W mech = ρVc p ΔT → ΔT =
η pump-motorW elect,in ρVc p
=
0.74 × (8785 kJ/s) (950.6 kg/m 3 )(1.5 m 3 /s)(4.229 kJ/kg ⋅ °C)
= 1.1°C
Therefore, the temperature of water will rise at least 1.1°C, which is more than the 0.5°C drop in temperature (in reality, the temperature rise will be more since the energy dissipation due to pump inefficiency will also appear as temperature rise of water). Thus we conclude that the frictional heating during flow can more than make up for the temperature drop caused by heat loss. Discussion The pumping power requirement and the associated cost can be reduced by using a larger diameter pipe. But the cost savings should be compared to the increased cost of larger diameter pipe.
8-61 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-86E Solution The air discharge rate of a clothes drier with no ducts is given. The flow rate when duct work is attached is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects in the duct are negligible, and thus the flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used. 4 The losses at the vent and its proximity are negligible. 5 The effect of the kinetic energy correction factor on discharge stream is negligible, α = 1. Properties The density of air at 1 atm and 120°F is ρ = 0.06843 lbm/ft3. The roughness of galvanized iron pipe is ε = 0.0005 ft. The loss coefficient is KL ≈ 0 for a well-rounded entrance with negligible loss, KL = 0.3 for a flanged 90° smooth bend, and KL = 1.0 for an exit. The friction factor of the duct is given to be 0.019. Analysis To determine the useful fan power input, we choose point 1 inside the drier sufficiently far from the vent, and point 2 at the exit on the same horizontal level so that z1 = z2 and P1 = P2, and the flow velocity at point 1 is negligible (V1 = 0) since it is far from the inlet of the fan. Also, the frictional piping losses between 1 and 2 are negligible, and the only loss involved is due to fan inefficiency. Then the energy equation for a control volume between 1 and 2 reduces to ⎛P ⎞ ⎛P ⎞ V2 V2 V2 m ⎜⎜ 1 + α 1 1 + gz1 ⎟⎟ + W fan = m ⎜⎜ 2 + α 2 2 + gz 2 ⎟⎟ + W turbine + E mech,loss → W fan, u = m 2 (1) 2 2 2 ⎝ρ ⎠ ⎝ ρ ⎠ and W = W − E . since α = 1 and E = E + E mech, loss
mech loss, fan
fan, u
mech loss, piping
fan
mech loss, fan
1.2 ft 3 /s V V The average velocity is V 2 = = = = 8.80 ft/s Ac πD 2 / 4 π (5 / 12 ft) 2 / 4 Now we attach the ductwork, and take point 3 to be at the duct exit so that the duct is included in the control volume. The energy equation for this control volume simplifies to V2 W fan, u = m 3 + m gh L Hot air (2) 2 3 Combining (1) and (2),
ρV2
V2 V 22 = ρV3 3 + ρV3 ghL → 2 2
V2 V2 V2 2 = V3 3 +V3 ghL 2 2
where V V3 V3 ft 3 /s V3 = 3 = = = 7.33V3 ft/s Ac πD 2 / 4 π (5 / 12 ft) 2 / 4
(3) Clothes drier
1 2 2 V2 15 ft ⎛ L ⎞V ⎛ ⎞V + + 3 × 0.3 + 1⎟ 3 = 2.58 3 hL = ⎜ f K L ⎟ 3 = ⎜ 0.019 5/12 ft 2g ⎝ D ⎠ 2g ⎝ ⎠ 2g Substituting into Eq. (3), V2 V2 (7.33V3 ) 2 (7.33V3 ) 2 V2 V2 2 = V3 3 +V3 g × 2.58 3 = V3 +V 3 × 2.58 = 96.2V33 2 2 2g 2 2 Solving for V and substituting the numerical values gives
∑
15 ft 5 in
2
3
⎛ V 22 ⎞ ⎟ V3 = ⎜ V2 ⎟ ⎜ 2 96 . 2 × ⎠ ⎝
1/ 3
⎛ 8.80 2 ⎞ ⎟ = ⎜⎜1.2 2 × 96.2 ⎟⎠ ⎝
1/ 3
= 0.78 ft 3 /s
Note that the flow rate decreased considerably for the same fan power input, as expected. We could also Discussion solve this problem by solving for the useful fan power first,
V2 (8.80 ft/s) 2 ⎛ 1 lbf 1W ⎞⎛ ⎞ = 0.13 W W fan, u = ρV2 2 = (0.06843 lbm/ft 3 )(1.2 ft 3 /s) ⎜ ⎟ 2 ⎟⎜ 2 2 ⋅ ft/s ⎠⎝ 0.737 lbf ⋅ ft/s ⎠ ⎝ 32.2 lbm Therefore, the fan supplies 0.13 W of useful mechanical power when the drier is running.
8-62 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-87 Solution Hot water in a water tank is circulated through a loop made of cast iron pipes at a specified average velocity. The required power input for the recirculating pump is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 Minor losses other than those for elbows and valves are negligible. Properties The density and dynamic viscosity of water at 60°C are ρ = 983.3 kg/m3, μ = 0.467×10-3 kg/m⋅s. The roughness of cast iron pipes is 0.00026 m. The loss coefficient is KL = 0.9 for a threaded 90° smooth bend and KL = 0.2 for a fully open gate valve. Analysis Since the water circulates continually and undergoes a cycle, we can take the entire recirculating system as the control volume, and choose points 1 and 2 at any location at the same point. Then the properties (pressure, elevation, and velocity) at 1 and 2 will be identical, and the energy equation will simplify to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g
→
h pump, u = h L
where 2
⎛ L ⎞V h L = h L,major + h L,minor = ⎜ f KL ⎟ 2 + ⎝ D ⎠ 2g since the diameter of the piping system is constant. Therefore, the pumping power is to be used to overcome the head losses in the flow. The flow rate and the Reynolds number are
∑
1.2 cm
40 m
Hot Water tank
V = VAc = V (πD 2 / 4) = (2.5 m/s)[π (0.012 m) 2 / 4] = 2.83 × 10 −4 m 3 /s Re =
ρVD (983.3 kg/m 3 )(2.5 m/s)(0.012 m) = = 63,200 μ 0.467 × 10 − 3 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 0.00026 m ε/D= = 0.0217 0.012 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 0.0217 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 Re f ⎟ ⎜ 3.7 f f 63200 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.05075. Then the total head loss, pressure drop, and the required pumping power input become 2 2 40 m ⎞ (2.5 m/s) ⎛ ⎞V ⎛ L = 55.8 m K L ⎟ 2 = ⎜ (0.05075) hL = ⎜ f + 6 × 0.9 + 2 × 0.2 ⎟ + 0.012 m ⎠ 2(9.81 m/s 2 ) ⎠ 2g ⎝ ⎝ D
∑
⎞⎛ 1 kPa ⎞ ⎛ 1 kN ⎟⎟⎜ = 538 kPa ΔP = ΔPL = ρgh L = (983.3 kg/m 3 )(9.81 m/s 2 )(55.8 m)⎜⎜ 2 ⎟ ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN/m ⎠ W elect =
W pump, u
η pump-motor
=
VΔP
η pump-motor
=
(2.83 × 10 -4 m 3 /s )(538 kPa ) ⎛ 1 kW ⎞ ⎟ = 0.217 kW ⎜ 0.70 ⎝ 1 kPa ⋅ m 3 /s ⎠
Therefore, the required power input of the recirculating pump is 217 W. Discussion It can be shown that the required pumping power input for the recirculating pump is 0.210 kW when the minor losses are not considered. Therefore, the minor losses can be neglected in this case without a major loss in accuracy.
8-63 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
8-88 Solution In the previous problem, the effect of average flow velocity on the power input to the recirculating pump for the same constant flow rate is to be investigated by varying the velocity from 0 to 3 m/s in increments of 0.3 m/s. Analysis
The EES Equations window is printed below, followed by the tabulated and plotted results.
g=9.81 rho=983.3 nu=mu/rho mu=0.000467 D=0.012 L=40 KL=6*0.9+2*0.2 Eff=0.7 Ac=pi*D^2/4 Vdot=V*Ac eps=0.00026 rf=eps/D "Reynolds number" Re=V*D/nu 1/sqrt(f)=-2*log10(rf/3.7+2.51/(Re*sqrt(f))) DP=(f*L/D+KL)*rho*V^2/2000 "kPa" W=Vdot*DP/Eff "kW" HL=(f*L/D+KL)*(V^2/(2*g))
0.4
V, m/s
Wpump, kW
ΔPL, kPa
Re
0.35
0.0 0.3 0.6 0.9 1.2 1.5 1.8 2.1 2.4 2.7 3.0
0.0000 0.0004 0.0031 0.0103 0.0243 0.0472 0.0814 0.1290 0.1922 0.2733 0.3746
0.0 8.3 32.0 71.0 125.3 195.0 279.9 380.1 495.7 626.6 772.8
0 7580 15160 22740 30320 37900 45480 53060 60640 68220 75800
0.3
W, kW
0.25 0.2 0.15 0.1 0.05 0 0
0.5
1
1.5
2
2.5
3
V, m/s Discussion As you might have suspected, the required power does not increase linearly with average velocity. Rather, the relationship is nearly quadratic. A larger diameter pipe would cut reduce the required pumping power considerably.
8-64 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-89 Solution Hot water in a water tank is circulated through a loop made of plastic pipes at a specified average velocity. The required power input for the recirculating pump is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The flow is fully developed. 3 The flow is turbulent so that the tabulated value of the loss coefficients can be used (to be verified). 4 Minor losses other than those for elbows and valves are negligible. Properties The density and viscosity of water at 60°C are ρ = 983.3 kg/m3, μ = 0.467×10-3 kg/m⋅s. Plastic pipes are smooth, and thus their roughness is very close to zero, ε = 0. The loss coefficient is KL = 0.9 for a threaded 90° smooth bend and KL = 0.2 for a fully open gate valve. Analysis Since the water circulates continually and undergoes a cycle, we can take the entire recirculating system as the control volume, and choose points 1 and 2 at any location at the same point. Then the properties (pressure, elevation, and velocity) at 1 and 2 will be identical, and the energy equation will simplify to P V2 P1 V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg ρg 2g 2g
→
h pump, u = h L
where 2
⎞V ⎛ L KL ⎟ 2 h L = h L,major + h L,minor = ⎜ f + ⎠ 2g ⎝ D since the diameter of the piping system is constant. Therefore, the pumping power is to be used to overcome the head losses in the flow. The flow rate and the Reynolds number are
∑
1.2 cm
40 m
Hot Water tank
V = VAc = V (πD 2 / 4) = (2.5 m/s)[π (0.012 m) 2 / 4] = 2.827 × 10 −4 m 3 /s Re =
ρVD (983.3 kg/m 3 )(2.5 m/s)(0.012 m) = = 63,200 μ 0.467 × 10 −3 kg/m ⋅ s
which is greater than 4000. Therefore, the flow is turbulent. The friction factor corresponding to the relative roughness of zero and this Reynolds number can simply be determined from the Moody chart. To avoid the reading error, we determine it from the Colebrook equation, ⎛ε / D ⎛ 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ 0 + ⎜ 3.7 Re f ⎟ ⎜ f f 63200 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.0198. Then the total head loss, pressure drop, and the required pumping power input become 2 2 40 m ⎞ (2.5 m/s) ⎛ ⎞V ⎛ L = 22.9 m K L ⎟ 2 = ⎜ (0.0198) hL = ⎜ f + 6 × 0.9 + 2 × 0.2 ⎟ + 0.012 m ⎠ 2(9.81 m/s 2 ) ⎠ 2g ⎝ ⎝ D
∑
⎞⎛ 1 kPa ⎞ ⎛ 1 kN ⎟⎟⎜ ΔP = ρgh L = (983.3 kg/m 3 )(9.81 m/s 2 )(22.9 m)⎜⎜ ⎟ = 221 kPa 2 1000 kg m/s ⋅ ⎠⎝ 1 kN/m ⎠ ⎝
W elect =
W pump, u
η pump-motor
=
VΔP
η pump-motor
=
(2.827 ×10 -4 m 3 /s)(221 kPa ) ⎛ 1 kW ⎞ ⎜ ⎟ = 0.0893 kW 0.70 ⎝ 1 kPa ⋅ m 3 /s ⎠
Therefore, the required power input of the recirculating pump is 89.3 W. Discussion It can be shown that the required pumping power input for the recirculating pump is 82.1 W when the minor losses are not considered. Therefore, the minor losses can be neglected in this case without a major loss in accuracy. Compared to the cast iron pipes of the previous problem, the plastic pipes reduced the required power by more than 50%, from 217 to 89.3 W. Furthermore, plastic pipes are lighter and easier to install, and they don’t rust.
8-65 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes Flow Rate and Velocity Measurements
8-90C Solution
We are to discus the primary considerations when choosing a flowmeter.
Analysis The primary considerations when selecting a flowmeter are cost, size, pressure drop, capacity, accuracy, and reliability. Discussion
As with just about everything you purchase, you usually get what you pay for.
8-91C Solution
We are to explain how a Pitot-static tube works and discuss its application.
Analysis A Pitot-static tube measures the difference between the stagnation and static pressure, which is the dynamic pressure, which is related to flow velocity by V = 2( P1 − P2 ) / ρ . Once the average flow velocity is determined, the flow rate is calculated from V = VA . The Pitot tube is inexpensive, highly reliable since it has no moving parts, it has c
very small pressure drop, and its accuracy (which is about 3%) is acceptable for most engineering applications. Discussion The term “Pitot tube” or “Pitot probe” is often used in place of “Pitot-static probe”. Technically, however, a Pitot probe measures only stagnation pressure, while a Pitot-static probe measures both stagnation and static pressures.
8-92C Solution
We are to discuss the operation of obstruction flowmeters.
Analysis An obstruction flowmeter measures the flow rate through a pipe by constricting the flow, and measuring the decrease in pressure due to the increase in velocity at (or downstream of) the constriction site. The flow rate for obstruction flowmeters is expressed as V = A C 2( P − P ) /[ ρ (1 − β 4 )] where A = π d 2 / 4 is the cross-sectional area of o 0
1
2
0
the obstruction and β = d / D is the ratio of obstruction diameter to the pipe diameter. Of the three types of obstruction flow meters, the orifice meter is the cheapest, smallest, and least accurate, and it causes the greatest head loss. The Venturi meter is the most expensive, the largest, the most accurate, and it causes the smallest head loss. The nozzle meter is between the orifice and Venturi meters in all aspects. Discussion As diameter ratio β decreases, the pressure drop across the flowmeter increases, leading to a larger minor head loss associated with the flowmeter, but increasing the sensitivity of the measurement.
8-93C Solution
We are to discuss the operation of positive displacement flowmeters.
Analysis A positive displacement flowmeter operates by trapping a certain amount of incoming fluid, displacing it to the discharge side of the meter, and counting the number of such discharge-recharge cycles to determine the total amount of fluid displaced. Positive displacement flowmeters are commonly used to meter gasoline, water, and natural gas because they are simple, reliable, inexpensive, and highly accurate even when the flow is unsteady. Discussion
In applications such as a gasoline meter, it is not the flow rate that is measured, but the flow volume.
8-66 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-94C Solution
We are to discuss the operation of a turbine flowmeter.
Analysis A turbine flowmeter consists of a cylindrical flow section that houses a turbine that is free to rotate, and a sensor that generates a pulse each time a marked point on the turbine passes by to determine the rate of rotation. Turbine flowmeters are relatively inexpensive, give highly accurate results (as accurate as 0.25%) over a wide range of flow rates, and cause a very small head loss. Discussion Turbine flowmeters must be calibrated so that a reading of the rpm of the turbine is translated into average velocity in the pipe or volume flow rate through the pipe.
8-95C Solution
We are to discuss the operation of rotameters.
Analysis A variable-area flowmeter (or rotameter) consists of a tapered conical transparent tube made of glass or plastic with a float inside that is free to move. As fluid flows through the tapered tube, the float rises within the tube to a location where the float weight, drag force, and buoyancy force balance each other. Variable-area flowmeters are very simple devices with no moving parts except for the float (but even the float remains stationary during steady operation), and thus they are very reliable. They are also very inexpensive, and they cause a relatively small head loss. Discussion There are also some disadvantages. For example, they must be mounted vertically, and most of them require a visual reading, and so cannot be automated or connected to a computer system.
8-96C Solution
We are to compare thermal and laser Doppler anemometers.
Analysis A thermal anemometer involves a very small electrically heated sensor (hot wire) which loses heat to the fluid, and the flow velocity is related to the electric current needed to maintain the sensor at a constant temperature. The flow velocity is determined by measuring the voltage applied or the electric current passing through the sensor. A laser Doppler anemometer (LDA) does not have a sensor that intrudes into flow. Instead, it uses two laser beams that intersect at the point where the flow velocity is to be measured, and it makes use of the frequency shift (the Doppler effect) due to fluid flow to measure velocity. Discussion Both of these devices measure the flow velocity at a point in the flow. Of the two, the hot wire system is much less expensive and has higher frequency resolution, but may interfere with the flow being measured.
8-97C Solution
We are to compare LDV and PIV.
Analysis Laser Doppler velocimetry (LDV) measures velocity at a point, but particle image velocimetry (PIV) provides velocity values simultaneously throughout an entire cross-section and thus it is a whole-field technique. PIV combines the accuracy of LDV with the capability of flow visualization, and provides instantaneous flow field mapping. Both methods are non-intrusive, and both utilize laser light beams. Discussion In both cases, optical access is required – a hot-wire system does not require optical access, but, like the LDV system, measures velocity only at a single point.
8-67 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-98 Solution The flow rate of ammonia is to be measured with flow nozzle equipped with a differential pressure gage. For a given pressure drop, the flow rate and the average flow velocity are to be determined. Assumptions
1 The flow is steady and incompressible. 2 The discharge coefficient of the nozzle meter is Cd = 0.96.
Properties respectively.
The density and dynamic viscosity of ammonia are given to be ρ = 624.6 kg/m3 and μ = 1.697×10-4 kg/m⋅s,
Analysis
The diameter ratio and the throat area of the meter are
β = d / D = 1.5 / 3 = 0.50
1.5 cm
3 cm
A0 = πd 2 / 4 = π (0.015 m) 2 / 4 = 1.767 × 10 − 4 m 2
Noting that ΔP = 4 kPa = 4000 N/m2, the flow rate becomes 2( P1 − P2 ) V = AoCd ρ (1 − β 4 )
ΔP = 4 kPa
⎛ 1 kg ⋅ m/s 2 ⎞ 2 × 4000 N/m ⎜ ⎟ = (1.767 × 10− 4 m 2 )(0.96) 3 4 ⎜ ⎟ (624.6 kg/m )((1 − 0.50 ) ⎝ 1 N ⎠ 2
= 0.627 × 10 − 3 m3 /s
which is equivalent to 0.627 L/s. The average flow velocity in the pipe is determined by dividing the flow rate by the crosssectional area of the pipe, V =
0.627 × 10−3 m 3 / s V V = = = 0.887 m/s 2 Ac πD / 4 π (0.03 m) 2 / 4
Discussion Re =
The Reynolds number of flow through the pipe is
ρVD (624.6 kg/m 3 )(0.887 m/s)(0.03 m) = = 9.79 × 10 4 μ 1.697 × 10 - 4 kg/m ⋅ s
Substituting the β and Re values into the orifice discharge coefficient relation gives 6.53β 0.5 6.53(0.50)0.5 Cd = 0.9975 −
Re0.5
= 0.9975 −
(9.79 × 104 )0.5
= 0.983
which is about 2% different than the assumed value of 0.96. Using this refined value of Cd, the flow rate becomes 0.642 L/s, which differs from our original result by only 2.4%. If the problem is solved using an equation solver such as EES, then the problem can be formulated using the curve-fit formula for Cd (which depends on Reynolds number), and all equations can be solved simultaneously by letting the equation solver perform the iterations as necessary.
8-68 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-99 Solution The flow rate of water through a circular pipe is to be determined by measuring the water velocity at several locations along a cross-section. For a given set of measurements, the flow rate is to be determined. Assumptions The points of measurements are sufficiently close so that the variation of velocity between points can be assumed to be linear.
The velocity measurements are given to be
Analysis
R, cm 0 1 2 3 4 5
V, m/s 6.4 6.1 5.2 4.4 2.0 0.0
•
1
•
2
•
3
•
4
•
5 cm r
The divide the cross-section of the pipe into 1-cm thick annual regions, as shown in the figure. Using midpoint velocity values for each section, the flow rate is determined to be V = VdA ≅ Vπ ( r 2 − r 2 )
∫
Ac
c
∑
out
in
⎛ 6.4 + 6.1 ⎞ ⎛ 6.1 + 5.2 ⎞ ⎛ 5.2 + 4.4 ⎞ 2 2 2 2 2 = π⎜ ⎟(0.01 − 0) + π ⎜ ⎟(0.02 − 0.01 ) + π ⎜ ⎟(0.03 − 0.02 ) 2 2 2 ⎝ ⎠ ⎝ ⎠ ⎝ ⎠ ⎛ 4.4 + 2.0 ⎞ ⎛ 2.0 + 0 ⎞ 2 2 2 2 +π⎜ ⎟(0.05 − 0.04 ) ⎟(0.04 − 0.02 ) + π ⎜ 2 2 ⎝ ⎠ ⎝ ⎠ = 0.0297 m 3 /s
Discussion We can also solve this problem by curve-fitting the given data using a second-degree polynomial, and then performing the integration.
8-69 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-100E Solution The flow rate of water is to be measured with an orifice meter. For a given pressure drop across the orifice plate, the flow rate, the average velocity, and the head loss are to be determined. 1 The flow is steady and incompressible. 2 The discharge coefficient of the orifice meter is Cd = 0.61.
Assumptions
Properties The density and dynamic viscosity of water are given to be ρ = 62.36 kg/m3 and μ = 7.536×10-4 lbm/ft⋅s, respectively. We take the density of mercury to be 847 lbm/ft3.
The diameter ratio and the throat area of the orifice are
Analysis
β = d / D = 2 / 4 = 0.50 4 in
A0 = πd 2 / 4 = π (2 / 12 ft) 2 / 4 = 0.02182 ft 2
2 in
The pressure drop across the orifice plate can be expressed as ΔP = P1 − P2 = ( ρ Hg − ρ f ) gh
6 in Then the flow rate relation for obstruction meters becomes 2( ρ Hg − ρf ) gh 2( ρ Hg / ρf − 1) gh 2( P1 − P2 ) V = AoCd = AoCd = AoCd 4 4 ρ (1 − β ) ρf (1 − β ) 1− β 4
Substituting, the flow rate is determined to be 2(847 / 62.36 − 1)(32.2 ft/s 2 )(6 / 12 ft ) = 0.277 ft 3 /s V = (0.02182 ft 2 )(0.61) 1 − 0.504
The average velocity in the pipe is determined by dividing the flow rate by the cross-sectional area of the pipe, V =
V V 0.277 ft 3 / s = = = 3.17 ft/s 2 Ac πD / 4 π (4 / 12 ft) 2 / 4
The percent pressure (or head) loss for orifice meters is given in Fig. 8-59 for β = 0.5 to be 74%. Therefore, noting that the density of mercury is 13.6 times that of water, hL = ( Permanent loss fraction )( Total head loss ) = 0.74 ( 0.50 ft Hg ) = 0.37 ft Hg = 5.03 ft H2 O
The head loss between the two measurement sections can also be estimated from the energy equation. Since z1 = z2, the head form of the energy equation simplifies to ⎡ ⎤ P − P V 2 − V12 ρ Hg ghHg ⎣( D / d ) − 1⎦ V1 = − hL ≈ 1 2 − 2 2g 2g ρf g ρf g 4
= SG Hg hHg
2
2 ⎡( D / d )4 − 1⎤ V12 ⎡ 24 − 1⎦⎤ ( 3.17 ft/s ) ⎣ ⎦ ⎣ − = 13.6 ( 0.50 ft ) − = 4.46 ft H2 O 2g 2 ( 32.2 ft/s 2 )
This head loss, though a reasonable estimate, is lower than the exact one calculated above because it does not take into account irreversible losses downstream of the pressure taps, where the flow is still “recovering,” and is not yet fully developed. Discussion
Re =
The Reynolds number of flow through the pipe is
ρVD (62.36 kg/m 3 )(3.17 ft/s)(4/12 ft) = = 8.744 × 10 4 -4 μ 7.536 ×10 lbm/ft ⋅ s
Substituting β and Re values into the orifice discharge coefficient relation 91.71β 2.5 Cd = 0.5959 + 0.0312 β 2.1 − 0.184 β 8 + 0.75 Re
gives Cd = 0.606, which is very close to the assumed value of 0.61. Using this refined value of Cd, the flow rate becomes 0.275 ft3/s, which differs from our original result by less than 1%. Therefore, it is convenient to analyze orifice meters using the recommended value of Cd = 0.61 for the discharge coefficient, and then to verify the assumed value.
8-70 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-101E Solution The flow rate of water is to be measured with an orifice meter. For a given pressure drop across the orifice plate, the flow rate, the average velocity, and the head loss are to be determined. 1 The flow is steady and incompressible. 2 The discharge coefficient of the orifice meter is Cd = 0.61.
Assumptions
Properties The density and dynamic viscosity of water are given to be ρ = 62.36 kg/m3 and μ = 7.536×10-4 lbm/ft⋅s, respectively. We take the density of mercury to be 847 lbm/ft3.
The diameter ratio and the throat area of the orifice are
Analysis
β = d / D = 2 / 4 = 0.50 A0 = πd 2 / 4 = π (2 / 12 ft) 2 / 4 = 0.02182 ft 2
4 in
2 in
The pressure drop across the orifice plate can be expressed as ΔP = P1 − P2 = ( ρ Hg − ρ f ) gh
Then the flow rate relation for obstruction meters becomes 2( ρ Hg − ρf ) gh 2( ρ Hg / ρf − 1) gh 2( P1 − P2 ) = AoCd = AoCd V = AoCd 4 4 ρ (1 − β ) ρf (1 − β ) 1− β 4
9 in
Substituting, the flow rate is determined to be 2(847 / 62.36 − 1)(32.2 ft/s 2 )(9 / 12 ft ) V = (0.02182 ft 2 )(0.61) = 0.339 ft 3 /s 1 − 0.504
The average velocity in the pipe is determined by dividing the flow rate by the cross-sectional area of the pipe, V =
V V 0.339 ft 3 / s = = = 3.88 ft/s Ac πD 2 / 4 π (4 / 12 ft) 2 / 4
The percent pressure (or head) loss for orifice meters is given in Fig. 8-59 for β = 0.5 to be 74%. Therefore, noting that the density of mercury is 13.6 times that of water, hL = (Permanent loss fraction )(Total head loss ) = 0.74(0.75 ft Hg) = 0.555 ft Hg = 7.55 ft H2O
The head loss between the two measurement sections can also be estimated from the energy equation. Since z1 = z2, the head form of the energy equation simplifies to ⎡ ⎤ P − P V 2 − V12 ρ Hg ghHg ⎣( D / d ) − 1⎦ V1 = − hL ≈ 1 2 − 2 ρf g ρf g 2g 2g 4
= SG Hg hHg
2
2 ⎡( D / d )4 − 1⎤ V12 ⎡ 24 − 1⎦⎤ ( 3.88 ft/s ) ⎣ ⎦ ⎣ − = 13.6 ( 0.75 ft ) − = 6.69 ft H2 O 2g 2 ( 32.2 ft/s 2 )
This head loss, though a reasonable estimate, is lower than the exact one calculated above because it does not take into account irreversible losses downstream of the pressure taps, where the flow is still “recovering,” and is not yet fully developed. Discussion Re =
The Reynolds number of flow through the pipe is
ρVD (62.36 kg/m 3 )(3.88 ft/s)(4/12 ft) = = 1.070 × 10 5 -4 μ 7.536 × 10 lbm/ft ⋅ s
Substituting β and Re values into the orifice discharge coefficient relation 91.71β 2.5 Cd = 0.5959 + 0.0312 β 2.1 − 0.184 β 8 + 0.75 Re
gives Cd = 0.605, which is very close to the assumed value of 0.61. Using this refined value of Cd, the flow rate becomes 0.336 ft3/s, which differs from our original result by less than 1%. Therefore, it is convenient to analyze orifice meters using the recommended value of Cd = 0.61 for the discharge coefficient, and then to verify the assumed value.
8-71 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-102 Solution The flow rate of water is measured with an orifice meter. The pressure difference indicated by the orifice meter and the head loss are to be determined. Assumptions
1 The flow is steady and incompressible. 2 The discharge coefficient of the orifice meter is Cd = 0.61.
Properties respectively.
The density and dynamic viscosity of water are given to be ρ = 998 kg/m3 and μ = 1.002×10-4 kg/m⋅s,
Analysis
The diameter ratio and the throat area of the orifice are
β = d / D = 30 / 50 = 0.60
50 cm
A0 = πd 2 / 4 = π (0.30 m) 2 / 4 = 0.07069 m 2
30 cm
For a pressure drop of ΔP = P1 − P2 across the orifice plate, the flow rate is expressed as 2( P1 − P2 ) V = AoCd ρ (1 − β 4 )
ΔP
Substituting, 0.25 m 3 / s = (0.07069 m 2 )(0.61)
2 ΔP (998 kg/m 3 )((1 − 0.60 4 )
which gives the pressure drop across the orifice plate to be ΔP = 14,600 kg ⋅ m/s 2 = 14.6 kPa
It corresponds to a water column height of ΔP 14,600 kg ⋅ m/s 2 = = 1.49 m ρ w g (998 kg/m 3 )(9.81 m/s 2 )
hw =
The percent pressure (or head) loss for orifice meters is given in Fig. 8-59 for β = 0.6 to be 64%. Therefore, hL = ( Permanent loss fraction )( Total head loss ) = 0.64 (1.49 m ) = 0.95 m H2 O
The head loss between the two measurement sections can also be estimated from the energy equation. Since z1 = z2, the head form of the energy equation simplifies to ⎡( D / d )4 − 1⎤ V12 ⎡( 50/30 )4 − 1⎤ (1.27 m/s )2 P1 − P2 V22 − V12 ⎣ ⎦ ⎦ − = hw − = 1.49 m − ⎣ = 0.940 m H2 O hL ≈ 2g 2g ρf g 2 ( 9.81 m/s 2 )
where
V1 =
V V 0.250 m3 / s = = = 1.27 m/s 2 Ac π D / 4 π ( 0.50 m )2 / 4
This head loss, though a reasonable estimate, is lower than the exact one calculated above because it does not take into account irreversible losses downstream of the pressure taps, where the flow is still “recovering,” and is not yet fully developed. Discussion Re =
The Reynolds number of flow through the pipe is
ρVD (998 kg/m 3 )(1.27 m/s)(0.50 m) = = 6.32 × 10 5 μ 1.002 × 10 -3 kg/m ⋅ s
Substituting β and Re values into the orifice discharge coefficient relation 91.71β 2.5 Cd = 0.5959 + 0.0312 β 2.1 − 0.184 β 8 + 0.75 Re
gives Cd = 0.605, which is very close to the assumed value of 0.61.
8-72 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-103 Solution A Venturi meter equipped with a differential pressure gage is used to measure to flow rate of water through a horizontal pipe. For a given pressure drop, the volume flow rate of water and the average velocity through the pipe are to be determined. Assumptions
The flow is steady and incompressible.
Properties The density of water is given to be ρ = 999.1 kg/m3. The discharge coefficient of Venturi meter is given to be Cd = 0.98. Analysis the meter are
3 cm
5 cm
The diameter ratio and the throat area of
β = d / D = 3 / 5 = 0.60 A0 = πd 2 / 4 = π (0.03 m) 2 / 4 = 7.069 × 10 − 4 m 2
ΔP
Noting that ΔP = 5 kPa = 5000 N/m , the flow rate becomes 2
Differential pressure meter
2( P1 − P2 ) V = AoCd ρ (1 − β 4 ) = (7.069 × 10 − 4 m 2 )(0.98)
⎛ 1 kg ⋅ m/s 2 ⎞ 2 × 5000 N/m 2 ⎜ ⎟ 3 4 ⎜ ⎟ 1N (999.1 kg/m )((1 − 0.60 ) ⎝ ⎠
= 0.00235 m3 /s
which is equivalent to 2.35 L/s. The average flow velocity in the pipe is determined by dividing the flow rate by the crosssectional area of the pipe, V =
Discussion
V V 0.00235 m 3 / s = = = 1.20 m/s 2 Ac πD / 4 π (0.05 m) 2 / 4
Note that the flow rate is proportional to the square root of pressure difference across the Venturi meter.
8-73 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
8-104 Solution The previous problem is reconsidered. The variation of flow rate as the pressure drop varies from 1 kPa to 10 kPa at intervals of 1 kPa is to be investigated, and the results are to be plotted. Analysis
The EES Equations window is printed below, followed by the tabulated and plotted results.
rho=999.1 "kg/m3" D=0.05 "m" d0=0.03 "m" beta=d0/D A0=pi*d0^2/4 Cd=0.98 Vol=A0*Cd*SQRT(2*DeltaP*1000/(rho*(1-beta^4)))*1000 "L/s" 3.5
Flow rate L/s 1.05 1.49 1.82 2.10 2.35 2.57 2.78 2.97 3.15 3.32
3
2.5
Flow rate, L/s
Pressure Drop ΔP, kPa 1 2 3 4 5 6 7 8 9 10
2
1.5
1 1
2
3
4
5
6
ΔP, kPa
7
8
9
10
Discussion This type of plot can be thought of as a calibration plot for the flowmeter, although a real calibration plot would use actual experimental data rather than data from equations. It would be interesting to compare the above plot to experimental data to see how close the predictions are.
8-74 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-105 Solution A Venturi meter equipped with a water manometer is used to measure to flow rate of air through a duct. For a specified maximum differential height for the manometer, the maximum mass flow rate of air that can be measured is to be determined. Assumptions
The flow is steady and incompressible.
Properties The density of air is given to be ρair = 1.204 kg/m3. We take the density of water to be ρw = 1000 kg/m3. The discharge coefficient of Venturi meter is given to be Cd = 0.98.
6 cm
15 cm
The diameter ratio and the throat area of the
Analysis meter are
h Water manometer
β = d / D = 6 / 15 = 0.40 A0 = πd 2 / 4 = π (0.06 m) 2 / 4 = 0.002827 m 2
The pressure drop across the Venturi meter can be expressed as ΔP = P1 − P2 = ( ρ w − ρ f ) gh
Then the flow rate relation for obstruction meters becomes 2( P1 − P2 ) 2( ρ w − ρ f ) gh 2( ρ w / ρair − 1) gh V = AoCd = AoCd 4 = AoCd 4 ρ (1 − β ) ρf (1 − β ) 1− β 4
Substituting and using h = 0.40 m, the maximum volume flow rate is determined to be 2(1000 / 1.204 − 1)(9.81 m/s 2 )(0.40 m) V = (0.002827 m 2 )(0.98) = 0.2265 m 3 / s 1 − 0.404
Then the maximum mass flow rate this Venturi meter can measure is m = ρV = (1.204 kg/m 3 )(0.2265 m 3/s ) = 0.273 kg/s
Also, the average flow velocity in the duct is V =
Discussion manometer.
V V 0.2265 m 3 / s = = = 12.8 m/s Ac πD 2 / 4 π (0.15 m) 2 / 4
Note that the maximum available differential height limits the flow rates that can be measured with a
8-75 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-106 Solution A Venturi meter equipped with a water manometer is used to measure to flow rate of air through a duct. For a specified maximum differential height for the manometer, the maximum mass flow rate of air that can be measured is to be determined. Assumptions
The flow is steady and incompressible.
Properties The density of air is given to be ρair = 1.204 kg/m3. We take the density of water to be ρw = 1000 kg/m3. The discharge coefficient of Venturi meter is given to be Cd = 0.98. Analysis The diameter ratio and the throat area of the meter are
7.5 cm
15 cm
h
β = d / D = 7.5 / 15 = 0.50
Water manometer
A0 = πd 2 / 4 = π (0.075 m) 2 / 4 = 0.004418 m 2
The pressure drop across the Venturi meter can be expressed as ΔP = P1 − P2 = ( ρ w − ρ f ) gh
Then the flow rate relation for obstruction meters becomes 2( P1 − P2 ) 2( ρ w − ρ f ) gh 2( ρ w / ρair − 1) gh = AoCd = AoCd V = AoCd ρ (1 − β 4 ) ρf (1 − β 4 ) 1− β 4
Substituting and using h = 0.40 m, the maximum volume flow rate is determined to be 2(1000 / 1.204 − 1)(9.81 m/s 2 )(0.40 m) V = (0.004418 m 2 )(0.98) = 0.3608 m 3 / s 1 − 0.504
Then the maximum mass flow rate this Venturi meter can measure is m = ρV = (1.204 kg/m 3 )(0.3608 m 3/s ) = 0.434 kg/s
Also, the average flow velocity in the duct is V =
Discussion manometer.
V V 0.3608 m 3 / s = = = 20.4 m/s 2 Ac πD / 4 π (0.15 m) 2 / 4
Note that the maximum available differential height limits the flow rates that can be measured with a
8-76 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-107 Solution A Venturi meter equipped with a differential pressure gage is used to measure the flow rate of liquid propane through a vertical pipe. For a given pressure drop, the volume flow rate is to be determined. Assumptions
The flow is steady and incompressible.
Properties The density of propane is given to be ρ = 514.7 kg/m3. The discharge coefficient of Venturi meter is given to be Cd = 0.98. Analysis
The diameter ratio and the throat area of the meter are
β = d / D = 5 / 8 = 0.625 A0 = πd 2 / 4 = π (0.05 m) 2 / 4 = 0.001963 m 2
Noting that ΔP = 7 kPa = 7000 N/m2, the flow rate becomes
5 cm
2( P1 − P2 ) V = AoCd ρ (1 − β 4 )
ΔP = 7 kPa
⎛ 1 kg ⋅ m/s 2 ⎞ 2 × 7000 N/m ⎜ ⎟ = (0.001963 m 2 )(0.98) 3 4 ⎜ ⎟ (514.7 kg/m )((1 − 0.625 ) ⎝ 1 N ⎠ 2
30 cm
= 0.0109 m3 /s
which is equivalent to 10.9 L/s. Also, the average flow velocity in the pipe is 0.0109 m 3 / s V V V = = = = 2.17 m/s Ac πD 2 / 4 π (0.08 m) 2 / 4
8 cm
Discussion Note that the elevation difference between the locations of the two probes does not enter the analysis since the pressure gage measures the pressure differential at a specified location. When there is no flow through the Venturi meter, for example, the pressure gage would read zero.
8-77 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-108 Solution The flow rate of water is to be measured with flow nozzle equipped with a differential pressure gage. For a given pressure drop, the flow rate, the average flow velocity, and head loss are to be determined. Assumptions
1 The flow is steady and incompressible. 2 The discharge coefficient of the nozzle meter is Cd = 0.96.
Properties respectively.
The density and dynamic viscosity of water are given to be ρ = 999.7 kg/m3 and μ = 1.307×10-3 kg/m⋅s,
Analysis
The diameter ratio and the throat area of the meter are
β = d / D = 1.5 / 3 = 0.50
1.5 cm
3 cm
A0 = πd 2 / 4 = π (0.015 m) 2 / 4 = 1.767 × 10 − 4 m 2
Noting that ΔP = 4 kPa = 4000 N/m2, the flow rate becomes 2( P1 − P2 ) V = AoCd ρ (1 − β 4 ) ⎛ 1 kg ⋅ m/s 2 ⎞ 2 × 3000 N/m 2 ⎜ ⎟ = (1.767 × 10− 4 m 2 )(0.96) ⎟ (999.7 kg/m 3 )((1 − 0.50 4 ) ⎜⎝ 1 N ⎠
ΔP = 3 kPa
= 0.429 × 10 − 3 m3 /s
which is equivalent to 0.429 L/s. The average flow velocity in the pipe is determined by dividing the flow rate by the crosssectional area of the pipe, V =
0.429 × 10−3 m 3 / s V V = = = 0.607 m/s Ac πD 2 / 4 π (0.03 m) 2 / 4
The water column height corresponding to a pressure drop of 3 kPa is ΔP 3000 kg ⋅ m/s 2 = = 0.306 m ρ w g (999.7 kg/m 3 )(9.81 m/s 2 )
hw =
The percent pressure (or head) loss for nozzle meters is given in Fig. 8-59 for β = 0.5 to be 62%. Therefore, hL = (Permanent loss fraction )(Total head loss) = 0.62(0.306 m) = 0.19 m H2O
The head loss between the two measurement sections can be determined from the energy equation, which simplifies to (for z1 = z2) hL =
Discussion Re =
P1 − P2 V22 − V12 [( D / d ) 4 − 1]V12 [(3/1.5) 4 − 1](0.607 m/s) 2 − = hw − = 0.306 m − = 0.024 m H2O 2g 2g ρf g 2(9.81 m/s 2 )
The Reynolds number of flow through the pipe is
ρVD (999.7 kg/m 3 )(0.607 m/s)(0.03 m) = = 1.39 × 10 4 μ 1.307 × 10 -3 kg/m ⋅ s
Substituting the β and Re values into the orifice discharge coefficient relation gives 6.53β 0.5 6.53(0.50)0.5 Cd = 0.9975 −
Re 0.5
= 0.9975 −
(1.39 × 104 )0.5
= 0.958
which is practically identical to the assumed value of 0.96.
8-78 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-109 Solution A kerosene tank is filled with a hose equipped with a nozzle meter. For a specified filling time, the pressure difference indicated by the nozzle meter is to be determined. Assumptions
1 The flow is steady and incompressible. 2 The discharge coefficient of the nozzle meter is Cd = 0.96.
Properties
The density of kerosene is given to be ρ = 820 kg/m3.
Analysis
The diameter ratio and the throat area of the meter are
β = d / D = 1.5 / 2 = 0.75
1.5 cm
2 cm
A0 = πd / 4 = π (0.015 m) / 4 = 1.767 × 10 2
2
−4
m
2
To fill a 16-L tank in 20 s, the flow rate must be V 16 L V = tank = = 0.8 L/s Δt 20 s
ΔP
For a pressure drop of ΔP = P1 − P2 across the meter, the flow rate is expressed as 2( P1 − P2 ) V = AoCd ρ (1 − β 4 )
Substituting, 0.0008 m 3 / s = (1.767 × 10− 4 m 2 )(0.96)
2ΔP (820 kg/m 3 )((1 − 0.754 )
which gives the pressure drop across the meter to be ΔP = 6230 kg ⋅ m/s 2 = 6.23 kPa
Discussion
Note that the flow rate is proportional to the square root of pressure difference across the nozzle meter.
8-79 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-110 Solution The flow rate of water is to be measured with flow nozzle equipped with an inverted air-water manometer. For a given differential height, the flow rate and head loss caused by the nozzle meter are to be determined. Assumptions
1 The flow is steady and incompressible. 2 The discharge coefficient of the nozzle meter is Cd = 0.96.
Properties respectively.
The density and dynamic viscosity of water are given to be ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s,
Analysis
The diameter ratio and the throat area of the meter are
β = d / D = 2 / 4 = 0.50 A0 = πd 2 / 4 = π (0.02 m) 2 / 4 = 3.142 × 10 − 4 m 2
32 cm
Noting that ΔP = 4 kPa = 4000 N/m2, the flow rate becomes 2( P1 − P2 ) = AoCd V = AoCd ρ (1 − β 4 ) = (3.142 × 10
−4
2 ρ w gh 2 gh = AoCd 1− β 4 ρ w (1 − β 4 )
Water
2 cm
4 cm
2(9.81 m/s 2 )(0.32 m) m )(0.96) 1 − 0.504 2
= 0.781 × 10 − 3 m3 /s
which is equivalent to 0.781 L/s. The average flow velocity in the pipe is V =
V V 0.781 × 10 −3 m 3 / s = = = 0.621 m / s Ac πD 2 / 4 π (0.04 m) 2 / 4
The percent pressure (or head) loss for nozzle meters is given in Fig. 8-59 for β = 0.5 to be 62%. Therefore, hL = (Permanent loss fraction )(Total head loss) = 0.62(0.32 m) = 0.20 m H2O
The head loss between the two measurement sections can be determined from the energy equation, which simplifies to (z1 =z2) 2 ⎡( D / d ) 4 − 1⎤ V12 ⎡ (4/2) 4 − 1⎦⎤ ( 0.621 m/s ) P1 − P2 V22 − V12 ⎦ hL = − = hw − ⎣ = 0.32 m − ⎣ = 0.025 m H 2 O 2g 2g ρf g 2 ( 9.81 m/s 2 )
Discussion
Re =
The Reynolds number of flow through the pipe is
ρVD (998 kg/m 3 )(0.621 m/s)(0.04 m) = = 2.47 × 10 4 μ 1.002 × 10 -3 kg/m ⋅ s
Substituting the β and Re values into the orifice discharge coefficient relation gives 6.53β 0.5 6.53(0.50)0.5 Cd = 0.9975 −
Re0.5
= 0.9975 −
(2.47 × 104 )0.5
= 0.968
which is almost identical to the assumed value of 0.96.
8-80 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-111E Solution A Venturi meter equipped with a differential pressure meter is used to measure to flow rate of refrigerant134a through a horizontal pipe. For a measured pressure drop, the volume flow rate is to be determined. Assumptions
The flow is steady and incompressible.
Properties The density of R-134a is given to be ρ = 83.31 lbm/ft3. The discharge coefficient of Venturi meter is given to be Cd = 0.98. Analysis the meter are
2 in
5 in
The diameter ratio and the throat area of
β = d / D = 2 / 5 = 0.40
ΔP = 7.4 psi
A0 = πd 2 / 4 = π (2 / 12 ft) 2 / 4 = 0.02182 ft 2
Noting that ΔP = 7.4 psi = 7.4×144 lbf/ft2, the flow rate becomes Differential pressure meter
2( P1 − P2 ) V = AoCd ρ (1 − β 4 ) = (0.02182 ft 2 )(0.98)
⎛ 32.2 lbm ⋅ ft/s 2 ⎞ 2 × 7.4 × 144 lbf/ft 2 ⎜ ⎟ 3 4 ⎜ ⎟ 1 lbf (83.31 lbm/ft )((1 − 0.40 ) ⎝ ⎠
= 0.622 ft 3 /s
Also, the average flow velocity in the pipe is determined by dividing the flow rate by the cross-sectional area of the pipe, V =
Discussion
V V 0.622 ft 3 / s = = = 4.56 ft/s Ac πD 2 / 4 π (5 / 12 ft)2 / 4
Note that the flow rate is proportional to the square root of pressure difference across the Venturi meter.
8-81 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes Review Problems
8-112 Solution A compressor takes in air at a specified rate at the outdoor conditions. The useful power used by the compressor to overcome the frictional losses in the duct is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 Air is an ideal gas. 4 The duct involves no components such as bends, valves, and connectors, and thus minor losses are negligible. 5 The flow section involves no work devices such as fans or turbines. Properties The properties of air at 1 atm = 101.3 kPa and 15°C are ρ0 = 1.225 kg/m3 and μ = 1.802×10-5 kg/m⋅s. The roughness of galvanized iron surfaces is ε = 0.00015 m. The dynamic viscosity is independent of pressure, but density of an ideal gas is proportional to pressure. The density of air at 95 kPa is 0.27 m3/s ρ = ( P / P0 ) ρ 0 = (95 / 101.3)(1.225 kg/m 3 ) = 1.149 kg/m 3 . 95 kPa Analysis The average velocity and the Reynolds number are
V=
V V 0.27 m 3 /s = = = 8.594 m/s Ac πD 2 / 4 π (0.20 m) 2 / 4
20 cm 8m
ρVD h (1.149 kg/m 3 )(8.594 m/s)(0.20 m) Re = = = 1.096 × 10 5 μ 1.802 × 10 −5 kg/m ⋅ s
Air compressor 150 hp
which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is 1.5 × 10 −4 m ε/D= = 7.5 × 10 − 4 0.20 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ ε / Dh 2.51 = −2.0 log⎜ + ⎜ 3.7 f Re f ⎝
1
⎞ ⎟ → ⎟ ⎠
⎛ 7.5 × 10 −4 2.51 = −2.0 log⎜ + ⎜ 3. 7 f 1.096 × 10 5 ⎝
1
⎞ ⎟ f ⎟⎠
It gives f = 0.02109. Then the pressure drop in the duct and the required pumping power become ΔP = ΔPL = f
L ρV 2 8 m (1.149 kg/m 3 )(8.594 m/s) 2 = 0.02109 D 2 0.20 m 2
⎛ 1N ⎜ ⎜ 1 kg ⋅ m/s 2 ⎝
⎞⎛ 1 Pa ⎞ ⎟⎜ ⎟⎝ 1 N/m 2 ⎟⎠ = 35.8 Pa ⎠
⎛ 1W ⎞ W pump, u = VΔP = (0.27 m 3 /s)(35.8 Pa )⎜ ⎟ = 9.66 W 3 1 Pa ⋅ m /s ⎝ ⎠ Discussion Note hat the pressure drop in the duct and the power needed to overcome it is very small (relative to 150 hp), and can be disregarded. The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.02086, which is very close to the Colebrook value. Also, the power input determined is the mechanical power that needs to be imparted to the fluid. The shaft power will be more than this due to fan inefficiency; the electrical power input will be even more due to motor inefficiency (but probably no more than 20 W).
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Chapter 8 Flow in Pipes 8-113 Solution Air enters the underwater section of a circular duct. The fan power needed to overcome the flow resistance in this section of the duct is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 Air is an ideal gas. 4 The duct involves no components such as bends, valves, and connectors. 5 The flow section involves no work devices such as fans or turbines. 6 The pressure of air is 1 atm. Properties The properties of air at 1 atm and 15°C are ρ0 = 1.225 kg/m3 and μ = -5 1.802×10 kg/m⋅s. The roughness of stainless steel pipes is ε = 0.000005 m.
Air, 3 m/s
The volume flow rate and the Reynolds number are
Analysis
V = VAc = V (πD 2 / 4 ) = (3 m/s)[π (0.20 m) 2 / 4] = 0.0942 m 3 /s
ρVD h (1.225 kg/m 3 )(3 m/s)(0.20 m) = = 4.079 × 10 4 μ 1.802 × 10 −5 kg/m ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Re =
Air River
5 × 10 −6 m ε/D= = 2.5 × 10 −5 0.20 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme),
⎛ ε / Dh ⎛ 2.5 × 10 −5 ⎞ 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + 4 ⎜ 3.7 ⎜ 3.7 f Re f ⎟⎠ f 4.079 × 10 f ⎟⎠ ⎝ ⎝ It gives f = 0.02195. Then the pressure drop in the duct and the required pumping power become 1
ΔP = ΔPL = f
W electric =
L ρV 2 15 m (1.225 kg/m 3 )(3 m/s) 2 = 0.02195 D 2 0.2 m 2
W pump, u
η pump - motor
=
VΔP
η pump - motor
=
⎛ 1N ⎜ ⎜ 1 kg ⋅ m/s 2 ⎝
⎞⎛ 1 Pa ⎞ ⎟⎜ ⎟⎝ 1 N/m 2 ⎟⎠ = 9.07 Pa ⎠
(0.0942 m 3 /s)(9.07 Pa ) ⎛ 1 W ⎞ =⎜ ⎟ = 1.4 W 0.62 ⎝ 1 Pa ⋅ m 3 /s ⎠
Discussion The friction factor could also be determined easily from the explicit Haaland relation. It would give f = 0.02175, which is sufficiently close to 0.02195. Assuming the pipe to be smooth would give 0.02187 for the friction factor, which is almost identical to the f value obtained from the Colebrook relation. Therefore, the duct can be treated as being smooth with negligible error.
8-114 Solution The velocity profile in fully developed laminar flow in a circular pipe is given. The radius of the pipe, the average velocity, and the maximum velocity are to be determined. Assumptions Analysis
The flow is steady, laminar, and fully developed.
The velocity profile in fully developed laminar flow in a circular pipe is ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠
The velocity profile in this case is given by u (r ) = 6(1 − 0.01r ) . Comparing the two relations above gives the pipe radius, the maximum velocity, and the average velocity to be 1 R2 = → R = 0.10 m 100
u(r)=umax(1-r2/R2)
R r
2
u max 6 m/s = = 3 m/s 2 2 In fully developed laminar pipe flow, average velocity is exactly half of maximum (centerline) velocity.
umax = 6 m/s Discussion
0
umax
→
V avg =
8-83 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-115E Solution The velocity profile in fully developed laminar flow in a circular pipe is given. The volume flow rate, the pressure drop, and the useful pumping power required to overcome this pressure drop are to be determined. Assumptions
1 The flow is steady, laminar, and fully developed. 2 The pipe is horizontal.
Properties The density and dynamic viscosity of water at 40°F are ρ = 62.42 lbm/ft3 and μ = 3.74 lbm/ft⋅h = -3 1.039×10 lbm/ft⋅s, respectively.
The velocity profile in fully developed laminar flow in a circular pipe is ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠ The velocity profile in this case is given by
Analysis
u(r)=umax(1-r2/R2)
R r
u (r ) = 0.8(1 − 625r 2 )
0
umax
Comparing the two relations above gives the pipe radius, the maximum velocity, and the average velocity to be 1 R2 = → R = 0.04 ft 625 umax = 0.8 ft/s V = V avg =
u max 0.8 ft/s = = 0.4 ft/s 2 2
Then the volume flow rate and the pressure drop become V = VAc = V (πR 2 ) = (0.4 ft/s)[π (0.04 ft) 2 ] = 0.00201 ft 3 /s
ΔPπD 4 Vhoriz = 128μL
→ 0.00201 ft 3 /s =
⎛ 32.2 lbm ⋅ ft/s 2 ⎜ 1 lbf lbm/ft ⋅ s)(80 ft) ⎜⎝
(ΔP )π (0.08 ft) 4 128(1.039 × 10
−3
⎞ ⎟ ⎟ ⎠
It gives ΔP = 5.16 lbf/ft 2 = 0.0358 psi
Then the useful pumping power requirement becomes
1W ⎛ ⎞ W pump, u = VΔP = (0.00201 ft 3 /s)(5.16 lbf/ft 2 )⎜ ⎟ = 0.014 W 0.737 lbf ⋅ ft/s ⎝ ⎠ Checking The flow was assumed to be laminar. To verify this assumption, we determine the Reynolds number: ρVD (62.42 lbm/ft 3 )(0.4 ft/s)(0.08 ft) Re = = = 1922 μ 1.039 × 10 − 3 lbm/ft ⋅ s
which is less than 2300. Therefore, the flow is laminar. Discussion Note that the pressure drop across the water pipe and the required power input to maintain flow is negligible. This is due to the very low flow velocity. Such water flows are the exception in practice rather than the rule.
8-84 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-116E Solution The velocity profile in fully developed laminar flow in a circular pipe is given. The volume flow rate, the pressure drop, and the useful pumping power required to overcome this pressure drop are to be determined. Assumptions
The flow is steady, laminar, and fully developed.
Properties The density and dynamic viscosity of water at 40°F are ρ = 62.42 lbm/ft3 and μ = 3.74 lbm/ft⋅h = -3 1.039×10 lbm/ft⋅s, respectively.
The velocity profile in fully developed laminar flow in a circular pipe is ⎛ r2 ⎞ u (r ) = u max ⎜⎜1 − 2 ⎟⎟ ⎝ R ⎠ The velocity profile in this case is given by
Analysis
u(r) = umax(1-r2/R2) R r 0
u ( r ) = 0.8(1 − 625r 2 )
umax
Comparing the two relations above gives the pipe radius, the maximum velocity, the average velocity, and the volume flow rate to be 1 R2 = → R = 0.04 ft 625 umax = 0.8 ft/s V = V avg =
u max 0.8 ft/s = = 0.4 ft/s 2 2
V = VAc = V (πR 2 ) = (0.4 ft/s)[π (0.04 ft) 2 ] = 0.00201 ft 3 /s
For uphill flow with an inclination of 12°, we have θ = +12°, and ⎛
⎞ ⎟ = 1038 lbf/ft 2 ⎝ 32.2 lbm ⋅ ft/s ⎠
ρgL sin θ = (62.42 lbm/ft 3 )(32.2 ft/s 2 )(80 ft ) sin 12°⎜
1 lbf
2
⎛ 32.2 lbm ⋅ ft/s 2 (ΔP − ρgL sin θ )πD 4 (ΔP − 1038)π (0.08 ft) 4 ⎜ → 0.00201 ft 3 /s = Vuphill = 3 − 128μL 1 lbf 128(1.039 × 10 lbm/ft ⋅ s)(80 ft) ⎜⎝
⎞ ⎟ ⎟ ⎠
It gives ΔP = 1043 lbf/ft 2 = 7.24 psi
Then the useful pumping power requirement becomes
1W ⎛ ⎞ W pump, u = VΔP = (0.00201 ft 3 /s)(1043 lbf/ft 2 )⎜ ⎟ = 2.84 W ⎝ 0.737 lbf ⋅ ft/s ⎠ Checking The flow was assumed to be laminar. To verify this assumption, we determine the Reynolds number: ρVD (62.42 lbm/ft 3 )(0.4 ft/s)(0.08 ft) Re = = = 1922 μ 1.039 × 10 − 3 lbm/ft ⋅ s
which is less than 2300. Therefore, the flow is laminar. Discussion Note that the pressure drop across the water pipe and the required power input to maintain flow is negligible. This is due to the very low flow velocity. Such water flows are the exception in practice rather than the rule.
8-85 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-117 Solution Water is discharged from a water reservoir through a circular pipe of diameter D at the side wall at a vertical distance H from the free surface with a reentrant section. A relation for the “equivalent diameter” of the reentrant pipe for use in relations for frictionless flow through a hole is to be obtained. Assumptions 1 The flow is steady and incompressible. 2 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 3 The water level in the reservoir remains constant. 4 The pipe is horizontal. 5 The entrance effects are negligible, and thus the flow is fully developed and the friction factor f is constant. 6 The effect of the kinetic energy correction factor is negligible, α = 1. Properties
The loss coefficient is KL = 0.8 for the reentrant section, and KL = 0 for the “frictionless” flow.
Analysis We take point 1 at the free surface of the reservoir and point 2 at the exit of the pipe, which is also taken as the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface of the reservoir is zero (V1 = 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L 2g 2g ρg ρg
→
H = α2
V 22 + hL 2g
where α2 = 1 and ⎛ L h L = h L , total = h L ,major + h L ,minor = ⎜ f + ⎝ D
∑K
L
2 2 ⎞ V2 ⎛ L ⎞V =⎜ f + KL ⎟ 2 ⎟ ⎠ 2g ⎝ D ⎠ 2g
since the diameter of the pipe is constant. Substituting and solving for V2 gives H=
2 V 22 ⎛ L ⎞V +⎜ f + KL ⎟ 2 2g ⎝ D ⎠ 2g
→
V2 =
1
2 gH 1 + fL / D + K L
H
πD V = AcV 2 = 4
2
2
D
Then the volume flow rate becomes 2 gH 1 + fL / D + K L
(1)
L
Note that in the special case of KL = 0 and f = 0 (frictionless flow), the velocity relation reduces to the Toricelli equation, V 2,frictionless = 2gz1 . The flow rate in this case through a hole of De (equivalent diameter) is V = Ac ,equiv V 2,frictionless =
2 πDequiv
2 gH (2) 4 Setting Eqs. (1) and (2) equal to each other gives the desired relation for the equivalent diameter, 2 πDequiv 2 gH πD 2 2 gH = 4 4 1 + fL / D + K L
which gives Dequiv =
D (1 + fL / D + K L ) 1 / 4
=
D (1 + 0.018 × 10 / 0.04 + 0.8) 1 / 4
= 0.63D = 0.025 m
Note that the effect of frictional losses of a pipe with a reentrant section is to reduce the diameter by about Discussion 40% in this case. Also, noting that the flow rate is proportional to the square of the diameter, we have 2 V ∝ Dequiv = (0.63D ) 2 = 0.40 D 2 . Therefore, the flow rate through a sharp-edged entrance is about two-thirds less compared to the frictionless flow case.
8-86 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-118 Solution A water tank open to the atmosphere is initially filled with water. The tank is drained to the atmosphere through a 90° horizontal bend of negligible length. The flow rate is to be determined for the cases of the bend being a flanged smooth bend and a miter bend without vanes. Assumptions 1 The flow is steady and incompressible. 2 The flow is turbulent so that the tabulated value of the loss coefficient can be used. 3 The water level in the tank remains constant. 4 The length of the bend and thus the frictional loss associated with its length is negligible. 5 The entrance is well-rounded, and the entrance loss is negligible. Properties
The loss coefficient is KL = 0.3 for a flanged smooth bend and KL = 1.1 for a miter bend without vanes.
Analysis (a) We take point 1 at the free surface of the tank, and point 2 at the exit of the bend, which is also taken as the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L 2g 2g ρg ρg
where the head loss is expressed as h L = K L z1 = α 2
V22 V2 + KL 2 2g 2g
→
→
z1 = α 2
V 22 + hL 2g
V2 . Substituting and solving for V2 gives 2g
2 gz1 = V22 ( α 2 + K L )
→
V2 =
2 gz1 α2 + KL
1
Then the flow rate becomes
πD V = ApipeV 2 = 4
2
5m
2 gz1
α2 + KL
3 cm
2
(a) Case 1 Flanged smooth bend (KL = 0.3):
πD V = AcV 2 = 4
2
2 gz1 π (0.03 m ) 2 = 4 α2 + KL
2(9.81 m/s 2 )(5 m) = 0.00603 m 3 /s = 6.03 L/s 1.05 + 0.3
(b) Case 2 Miter bend without vanes (KL = 1.1):
πD V = AcV 2 = 4
2
2 gz1 π (0.03 m ) 2 = 4 α2 + KL
2(9.81 m/s 2 )(5 m) = 0.00478 m 3 /s = 4.78 L/s 1.05 + 1.1
Discussion Note that the type of bend used has a significant effect on the flow rate, and a conscious effort should be made when selecting components in a piping system. If the effect of the kinetic energy correction factor is neglected, α2 = 1 and the flow rates become
(a) Case 1 (KL = 0.3):
πD V = AcV 2 = 4
(b) Case 2 (KL = 1.1):
πD V = AcV 2 = 4
2
2
2 gz1 π (0.03 m ) 2 = 1+ K L 4
2(9.81 m/s 2 )(5 m) = 0.00614 m 3 /s 1 + 0.3
2 gz1 π (0.03 m ) 2 = 1+ K L 4
2(9.81 m/s 2 )(5 m) = 0.00483 m 3 /s 1 + 1.1
Therefore, the effect of the kinetic energy correction factor is (6.14 - 6.03)/6.03 = 1.8% and (4.83 - 4.78)/4.83 = 1.0%, which is negligible.
8-87 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-119 [Also solved using EES on enclosed DVD]
Solution The piping system of a geothermal district heating system is being designed. The pipe diameter that will optimize the initial system cost and the energy cost is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses are negligible because of the large length-to-diameter ratio and the relatively small number of components that cause minor losses, the only significant energy loss arises from pipe friction. 4 The piping system is horizontal. 5 The properties of geothermal water are the same as fresh water. 6 The friction factor is constant at the given value. 7 The interest rate, the inflation rate, and the salvage value of the system are all zero. 8 The flow rate through the system remains constant. Properties
We take the density of water to be ρ = 1000 kg/m3. The friction factor is given to be f = 0.015.
Analysis The system operates in a loop, and thus we can take any point in the system as points 1 and 2 (the same point), and thus z1 = z2, V1 = V2, and P1 = P2. Then the energy equation for this piping system simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 1 2 + z 2 + h turbine, e + h L → hpump, u = h L 2g 2g ρg ρg That is, the pumping power is to be used to overcome the head losses due to friction in flow. When the minor losses are disregarded, the head loss for fully developed flow in a pipe of length L and diameter D can be expressed as ΔP = f
L ρV 2 D 2
Water
D
10,000 kg/s
The flow rate of geothermal water is m 10,000 kg/s V = = = 10 m 3 /s ρ 1000 kg/m 3 To expose the dependence of pressure drop on diameter, we express it in terms of the flow rate as ΔP = f
L = 10 km
2 16 L ρV 2 8 L ρV 2 L ρV 2 L ρ ⎛ V ⎞ ⎜ ⎟ = f f = f = D 2π 2 D 4 D 2 D 2 ⎜⎝ πD 2 / 4 ⎟⎠ D5 π 2
Then the required pumping power can be expressed as W pump, u ρV 3 VΔP V 8L ρV 2 8L W pump = = = f 5 = f η pump -motor η pump - motor η pump - motor D π 2 D 5 η pump - motor π 2 Note that the pumping power requirement is proportional to f and L, consistent with our intuitive expectation. Perhaps not so obvious is that power is proportional to the cube of flow rate. The fact that the power is inversely proportional to pipe diameter D to the fifth power averages that a slight increase in pipe diameter will manifest as a tremendous reduction in power dissipation due to friction in a long pipeline. Substituting the given values and expressing the diameter D in meters, 8(10,000 m) (1000 kg/m 3 )(10 m 3 /s) 3 W pump = (0.015) π 2 (0.80) D5 m5
⎛ 1 kN ⎜ ⎜ 1000 kg ⋅ m/s 2 ⎝
⎞⎛ 1 kW ⎞ 1.52 × 10 5 ⎟⎜ kW = ⎟⎝ 1 kN ⋅ m/s ⎟⎠ D5 ⎠
The number of hours in one year are 24×365 = 8760 h. Then the total amount of electric power used and its cost per year are 1.52 × 10 5 1.332 × 10 9 = E pump = W pump Δt = ( 8760 h) kWh/yr D5 D5 ⎛ 1.332 × 10 9 ⎞ 7.99 × 10 7 ⎟($0.06/kWh) = Energy cost = E pump × Unit cost = ⎜⎜ kWh/y $/yr ⎟ D5 D5 ⎝ ⎠ The installation cost of the system with a 30-year lifetime is given to be Cost = $106 D2 where D is in meters. The annual cost of the system is then 1/30th of it, which is Total cost $10 6 D 2 System cost = = = $3.33 × 10 4 D 2 (per year) Life time 30 yr Then the total annual cost of the system (installation + operating) becomes
8-88 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 7.99 × 10
7
+ 3.33 × 10 4 D 2 $/yr D5 The optimum pipe diameter is the value that minimizes this total, and it is determined by taking the derivative of the total cost with respect to D and setting it equal to zero,
Total cost = Energy cost + System cost =
∂(Total cost) 7.99 × 10 7 = −5 × + 2 × 3.33 × 10 4 D = 0 6 ∂D D Simplifying gives D7 = 5998 whose solution is D = 3.5 m This is the optimum pipe diameter that minimizes the total cost of the system under stated assumptions. A larger diameter pipe will increase the system cost more than it decreases the energy cost, and a smaller diameter pipe will increase the system cost more than it decreases the energy cost. Discussion The assumptions of zero interest and zero inflation are not realistic, and an actual economic analysis must consider these factors as they have a major effect on the pipe diameter. This is done by considering the time value of money, and expressing all the costs at the same time. Pipe purchase is a present cost, and energy expenditures are future annual costs spread out over the project lifetime. Thus, to provide consistent dollar comparisons between initial and future costs, all future energy costs must be expressed as a single present lump sum to reflect the time-value of money. Then we can compare pipe and energy costs on a consistent basis. Economists call the necessary factor the “Annuity Present Value Factor”, F. If interest rate is 10% per year with n = 30 years, then F = 9.427. Thus, if power costs $1,000,000/year for the next 30 years, then the present value of those future payments is $9,427,000 (and not $30,000,000!) if money is worth 10%. Alternatively, if you must pay $1,000,000 every year for 30 years, and you can today invest $9,437,000 at 10%, then you can meet 30 years of payments at the end of each year. The energy cost in this case can be determined by dividing the energy cost expression developed above by 9.427. This will result in a pipe diameter of D = 2.5 m. In an actual design, we also need to calculate the average flow velocity and the pressure head to make sure that they are reasonable. For a pipe diameter of 2.5 m, for example, the average flow velocity is 1.47 m/s and the pump pressure head is 5.6 m.
8-89 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-120 Solution Water is drained from a large reservoir through two pipes connected in series at a specified rate using a pump. The required pumping head and the minimum pumping power are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The pipes are horizontal. 3 The entrance effects are negligible, and thus the flow is fully developed. 4 The flow is turbulent so that the tabulated value of the loss coefficients can be used. 5 The pipes involve no components such as bends, valves, and other connectors that cause additional minor losses. 6 The reservoir is open to the atmosphere so that the pressure is atmospheric pressure at the free surface. 7 The water level in the reservoir remains constant. 8 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 15°C are ρ = 999.1 kg/m3 and μ = 1.138×10-3 kg/m⋅s. The loss coefficient is KL = 0.5 for a sharp-edged entrance. The roughness of cast iron pipes is ε = 0.00026 m. Analysis We take point 1 at the free surface of the tank, and point 2 and the reference level at the centerline of the pipe (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface of the tank is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to
P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + hturbine, e + h L ρg 2g ρg 2g where α2 = 1 and 2 ⎛ L ⎞V + h L = h L, total = h L,major + h L,minor = KL ⎟ ⎜f ⎝ D ⎠ 2g
∑
→
z1 + hpump, u = α 2
V 22 + hL 2g
∑
and the summation is over two pipes. Noting that the two pipes are connected in series and thus the flow rate through each of them is the same, the head loss for each pipe is determined as follows (we designate the first pipe by 1 and the second one by 2): 0.018 m 3 /s V V Pipe 1: V1 = = = = 6.366 m/s 1 Ac1 πD12 / 4 π (0.06 m) 2 / 4
ρV1 D1 (999.1 kg/m 3 )(6.366 m/s)(0.06 m) = = 335,300 μ 1.138 × 10 −3 kg/m ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Re 1 =
Water tank
30 m
10 L/s
2 35 m
20 m 0.00026 m Pump = 0.00433 0.06 m The friction factor corresponding to this relative roughness and the Reynolds number is, from the Colebrook equation, ⎛ 0.00433 ⎞ ⎛ ε / D1 1 2.51 ⎞⎟ 1 2.51 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 ⎟ ⎜ 3.7 ⎟ f1 f f f Re 335 , 300 1 1 ⎠ 1 1 ⎠ ⎝ ⎝ It gives f1 = 0.02941. The only minor loss is the entrance loss, which is KL = 0.5. Then the total head loss of the first pipe becomes 2 ⎛ L ⎞V 2 ⎛ 20 m ⎞ (6.366 m/s) h L1 = ⎜⎜ f 1 1 + K L ⎟⎟ 1 = ⎜ (0.02941) + 0.5 ⎟ = 21.3 m 0.06 m ⎠ 2(9.81 m/s 2 ) ⎝ D1 ⎠ 2g ⎝ 0.01 8m 3 /s V V = = = 14.32 m/s Pipe 2: V 2 = Ac 2 πD 22 / 4 π (0.04 m) 2 / 4
ε / D1 =
∑
ρV 2 D 2 (999.1 kg/m 3 )(14.32 m/s)(0.04 m) = = 502,900 μ 1.138 × 10 −3 kg/m ⋅ s which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Re 2 =
0.00026 m = 0.0065 0.04 m The friction factor corresponding to this relative roughness and the Reynolds number is, from the Colebrook equation,
ε / D2 =
8-90 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes ⎛ε / D ⎛ 0.0065 ⎞ 2.51 ⎞⎟ 1 2.51 2 ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + ⎜ 3.7 ⎟ ⎜ 3 .7 ⎟ f2 f f f Re 502 , 900 2 2 ⎠ 2 2 ⎠ ⎝ ⎝ It gives f2 = 0.03309. The second pipe involves no minor losses. Note that we do not consider the exit loss unless the exit velocity is dissipated within the system considered (in this case it is not). Then the head loss for the second pipe becomes L V2 35 m (14.32 m/s) 2 h L 2 = f 2 2 2 = (0.03309) = 302.6 m D2 2 g 0.04 m 2(9.81 m/s 2 ) The total head loss for two pipes connected in series is the sum of the head losses of the two pipes, 1
h L = h L , total = h L1 + h L 2 = 21.3 + 302.6 = 323.9 m
Then the pumping head and the minimum pumping power required (the pumping power in the absence of any inefficiencies of the pump and the motor, which is equivalent to the useful pumping power) become V2 (14.32 m/s) 2 h pump, u = α 2 2 + h L − z1 = (1) + 323.9 − 30 = 304.4 m 2g 2(9.81 m/s 2 ) W pump, u = VΔP = ρVghpump, u ⎛ ⎞⎛ 1 kW ⎞ 1 kN ⎟ = (999.1 kg/m 3 )(0.018 m 3 /s )(9.81 m/s 2 )(304.4 m)⎜⎜ ⎟ = 53.7 kW 2 ⎟⎜ ⎝ 1000 kg ⋅ m/s ⎠⎝ 1 kN ⋅ m/s ⎠ Therefore, the pump must supply a minimum of 53.7 kW of useful mechanical energy to water. Discussion Note that the shaft power of the pump must be greater than this to account for the pump inefficiency, and the electrical power supplied must be even greater to account for the motor inefficiency.
8-91 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
8-121
Solution In the previous problem, the effect of second pipe diameter on required pumping head for the same flow rate is to be investigated by varying the pipe diameter from 1 cm to 10 cm in increments of 1 cm. The EES Equations window is printed below, followed by the tabulated and plotted results.
rho=999.1 mu=0.001138 nu=mu/rho Vdot=0.018 "m3/s" g=9.81 "m/s2" z1=30 "m" L1=20 "m" D1=0.06 "m" Ac1=pi*D1^2/4 Re1=V1*D1/nu V1=Vdot/Ac1 eps1=0.00026 rf1=eps1/D1 1/sqrt(f1)=-2*log10(rf1/3.7+2.51/(Re1*sqrt(f1))) KL1=0.5 HL1=(f1*L1/D1+KL1)*V1^2/(2*g) L2=35 Re2=V2*D2/nu V2=Vdot/(pi*D2^2/4) eps2=0.00026 rf2=eps2/D2 1/sqrt(f2)=-2*log10(rf2/3.7+2.51/(Re2*sqrt(f2))) HL2=f2*(L2/D2)*V2^2/(2*g) HL=HL1+HL2 hpump=V2^2/(2*g)+HL-z1 Wpump=rho*Vdot*g*hpump/1000 "kW"
D2, m 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10
90000 80000 70000 60000
Wpump, kW
Analysis
50000 40000 30000 20000 10000 0 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09
0.1
D2, m
Wpump, kW 89632.5 2174.7 250.8 53.7 15.6 5.1 1.4 -0.0 -0.7 -1.1
hL2, m 505391.6 12168.0 1397.1 302.8 92.8 35.4 15.7 7.8 4.2 2.4
Re 2.012E+06 1.006E+06 6.707E+05 5.030E+05 4.024E+05 3.353E+05 2.874E+05 2.515E+05 2.236E+05 2.012E+05
Discussion Clearly, the power decreases quite rapidly with increasing diameter. This is not surprising, since the irreversible frictional head loss (major head loss) decreases significantly with increasing pipe diameter.
8-92 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-122
Solution Two pipes of identical diameter and material are connected in parallel. The length of one of the pipes is twice the length of the other. The ratio of the flow rates in the two pipes is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The flow is fully turbulent in both pipes and thus the friction factor is independent of the Reynolds number (it is the same for both pipes since they have the same material and diameter). 3 The minor losses are negligible. Analysis When two pipes are parallel in a piping system, the head loss for each pipe must be same. When the minor losses are disregarded, the head loss for fully developed flow in a pipe of length L and diameter D can be expressed as hL = f
2 2 L V 2 L 1 V 2 L V2 L 1 ⎛ V ⎞ L 1 ⎛ V ⎞ ⎜ ⎟ = f ⎜ ⎟ =8f = f = 8 f D g π 2D4 D 2g D 2 g ⎜⎝ Ac ⎟⎠ D 2 g ⎜⎝ πD 2 / 4 ⎟⎠ gπ 2 D 5
Solving for the flow rate gives
π 2 h L gD 5 k V = = 8 fL L
(k is a constant)
When the pipe diameter, friction factor, and the head loss is constant, which is the case here for parallel connection, the flow rate becomes inversely proportional to the square root of length L. Therefore, when the length is doubled, the flow rate will decrease by a factor of 20.5 = 1.41 since If
Then
VA =
VB =
k
D
LA
k LB
2L
A
=
k 2L A
=
k 2 LA
=
VA 2
= 0.707VA
Therefore, the ratio of the flow rates in the two pipes is 0.707.
B
L
D
Discussion Even though one pipe is twice as long as the other, the volume flow rate in the shorter pipe is not twice as much – the relationship is nonlinear.
8-93 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-123 [Also solved using EES on enclosed DVD]
Solution A pipeline that transports oil at a specified rate branches out into two parallel pipes made of commercial steel that reconnects downstream. The flow rates through each of the parallel pipes are to be determined. Assumptions 1 The flow is steady and incompressible. 2 Entrance effects are negligible, and thus the flow is fully developed. 3 Minor losses are disregarded. 4 Flows through both pipes are turbulent (to be verified). Properties The density and dynamic viscosity of oil at 40°C are ρ = 876 kg/m3 and μ = 0.2177 kg/m⋅s. The roughness of commercial steel pipes is ε = 0.000045 m. Analysis This problem cannot be solved directly since the velocities (or flow rates) in the pipes are not known. Below we will set up the equations to be solved by an equation solver. The head loss in two parallel branches must be the same, and the total flow rate must the sum of the flow rates in the parallel branches. Therefore, h L,1 = h L , 2
(1)
V = V1 + V2 → V1 + V2 = 3 (2) We designate the 30-cm diameter pipe by 1 and the 45-cm diameter pipe by 2. The average velocity, the relative roughness, the Reynolds number, friction factor, and the head loss in each pipe are expressed as V V1 V1 → = (3) V V1 = 1 = 1 Ac,1 πD12 / 4 π (0.30m) 2 / 4 V V2 V2 → V = (4) V2 = 2 = 2 1 Ac , 2 πD 22 / 4 π (0.45m) 2 / 4 30 cm 500 m rf 1 =
ε1 D1
rf 2 =
ε2 D2
= =
4.5 × 10 −5 m = 1.5 × 10 − 4 0.30 m
A 3 m3/s
−5
4.5 × 10 m = 1× 10 − 4 0.45 m
Re 1 =
ρV1 D1 μ
Re 2 =
ρV 2 D 2 μ
→ Re 1 =
B 800 m
3
(876 kg/m )V1 (0.30 m) 0.2177 kg/m ⋅ s
→ Re 2 =
⎛ε / D 2.51 1 = −2.0 log⎜ + ⎜ 3.7 f1 Re 1 f 1 ⎝ 1 f2
⎞ ⎟ → ⎟ ⎠
⎛ε / D 2.51 2 = −2.0 log⎜ + ⎜ 3.7 Re 2 f 2 ⎝ h L ,1 = f 1
L1 V12 D1 2 g
→
L 2 V 22 D2 2 g
→
45 cm
(5)
(876 kg/m 3 )V2 (0.45 m) 0.2177 kg/m ⋅ s
1
2
(6) ⎛ 1.5 × 10 − 4 2.51 = −2.0 log⎜ + ⎜ 3.7 f1 Re 1 f 1 ⎝
1
−4 ⎞ ⎛ ⎟ → 1 = −2.0 log⎜ 1× 10 + 2.51 ⎟ ⎜ 3.7 f2 Re 2 f 2 ⎠ ⎝
h L,1 = f 1
V12 500 m 0.30 m 2(9.81 m/s 2 )
⎞ ⎟ (7) ⎟ ⎠ ⎞ ⎟ (8) ⎟ ⎠
(9)
2
V2 800 m (10) 0.45 m 2(9.81 m/s 2 ) This is a system of 10 equations in 10 unknowns, and solving them simultaneously by an equation solver gives V = 0.91 m 3 /s, V = 2.09 m 3 /s , h L ,2 = f 2
1
h L,2 = f 2
2
V1 = 12.9 m/s, V2 = 13.1 m/s, h L,1 = h L,2 = 392 m Re1 = 15,540, Re2 = 23,800, f1 = 0.02785, f2 = 0.02505 Note that Re > 4000 for both pipes, and thus the assumption of turbulent flow is verified. Discussion This problem can also be solved by using an iterative approach, but it would be very time consuming. Equation solvers such as EES are invaluable for theses kinds of problems.
8-94 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-124 Solution The piping of a district heating system that transports hot water at a specified rate branches out into two parallel pipes made of commercial steel that reconnects downstream. The flow rates through each of the parallel pipes are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed. 3 The minor losses are given to be negligible. 4 Flows through both pipes are turbulent (to be verified). Properties The density and dynamic viscosity of water at 100°C are ρ = 957.9 kg/m3 and μ = 0.282×10-3 kg/m⋅s. The roughness of commercial steel pipes is ε = 0.000045 m. Analysis This problem cannot be solved directly since the velocities (or flow rates) in the pipes are not known. Below we will set up the equations to be solved by an equation solver. The head loss in two parallel branches must be the same, and the total flow rate must the sum of the flow rates in the parallel branches. Therefore, h L,1 = h L , 2
(1)
V = V1 + V2 → V1 + V2 = 3 (2) We designate the 30-cm diameter pipe by 1 and the 45-cm diameter pipe by 2. The average velocity, the relative roughness, the Reynolds number, friction factor, and the head loss in each pipe are expressed as V V1 V1 → V1 = (3) V1 = 1 = 2 Ac,1 πD1 / 4 π (0.30m) 2 / 4 V V2 V2 1 → V = (4) V2 = 2 = 2 30 cm 500 m Ac , 2 πD 22 / 4 π (0.45m) 2 / 4 rf 1 =
ε1 D1
rf 2 =
ε2 D2
=
4.5 × 10 −5 m = 1.5 × 10 − 4 0.30 m
=
4.5 × 10 −5 m = 1× 10 − 4 0.45 m
Re 1 =
ρV1 D1 μ
Re 2 =
ρV 2 D 2 μ
→ Re 1 =
A
3
3 m /s B 800 m
(957.9 kg/m 3 )V1 (0.30 m)
→ Re 2 =
(957.9 kg/m 3 )V 2 (0.45 m) 0.282 × 10 − 3 kg/m ⋅ s
⎛ε / D 2.51 1 = −2.0 log⎜ + ⎜ 3 . 7 f1 Re 1 f 1 ⎝ 1 f2
⎞ ⎟ → ⎟ ⎠
⎛ε / D 2.51 2 = −2.0 log⎜ + ⎜ 3.7 Re 2 f 2 ⎝ h L ,1 = f 1
L1 V12 D1 2 g
→
L 2 V 22 D2 2 g
→
45 cm
(5)
0.282 × 10 − 3 kg/m ⋅ s
1
2
(6)
⎛ 1.5 × 10 − 4 2.51 = −2.0 log⎜ + ⎜ 3 . 7 f1 Re 1 f 1 ⎝
1
−4 ⎞ ⎛ ⎟ → 1 = −2.0 log⎜ 1× 10 + 2.51 ⎟ ⎜ 3.7 f2 Re 2 f 2 ⎠ ⎝
h L,1 = f 1
V12 500 m 0.30 m 2(9.81 m/s 2 )
⎞ ⎟ (7) ⎟ ⎠ ⎞ ⎟ (8) ⎟ ⎠
(9)
2
V2 800 m (10) 0.45 m 2(9.81 m/s 2 ) This is a system of 10 equations in 10 unknowns, and their simultaneous solution by an equation solver gives V = 0.919 m 3 /s, V = 2.08 m 3 /s , h L ,2 = f 2
1
h L,2 = f 2
2
V1 = 13.0 m/s, V2 = 13.1 m/s, h L ,1 = h L,2 = 187 m Re1 = 1.324×107, Re2 = 2.00×107, f1 = 0.0131, f2 = 0.0121 Note that Re > 4000 for both pipes, and thus the assumption of turbulent flow is verified. Discussion This problem can also be solved by using a trial-and-error approach, but it will be very time consuming. Equation solvers such as EES are invaluable for these kinds of problems.
8-95 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-125E Solution A water fountain is to be installed at a remote location by attaching a cast iron pipe directly to a water main. For a specified flow rate, the minimum diameter of the piping system is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed and the friction factor is constant for the entire pipe. 3 The pressure at the water main remains constant. 4 There are no dynamic pressure effects at the pipe-water main connection, and the pressure at the pipe entrance is 60 psia. 5 Elevation difference between the pipe and the fountain is negligible (z2 = z1) . 6 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The roughness of cast iron pipe is ε = 0.00085 ft. The minor loss coefficient is KL = 0.5 for a sharpedged entrance, KL = 1.1 for a 90° miter bend without vanes, KL = 0.2 for a fully open gate valve, and KL = 5 for an angle valve. Analysis We choose point 1 in the water main near the entrance where the pressure is 60 psig and the velocity in the pipe to be low. We also take point 1 as the reference level. We take point 2 at the exit of the water fountain where the pressure is the atmospheric pressure (P2 = Patm) and the velocity is the discharge velocity. The energy equation for a control volume between these two points is P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L 2g 2g ρg ρg
→
P1,gage
ρg
=α2
V 22 + hL 2g
2 ⎞ V2
⎛ L KL ⎟ where α2 = 1 and h L = h L ,total = h L,major + h L ,minor = ⎜ f + ⎝ D ⎠ 2g since the diameter of the piping system is constant. Then the energy equation becomes
∑
⎛ 144 lbf/ft 2 ⎞⎛ 32.2 lbm ⋅ ft/s 2 ⎜ ⎟⎜ 1 lbf (62.3 lbm/ft 3 )(32.2 ft/s 2 ) ⎜⎝ 1 psi ⎟⎠⎜⎝ The average velocity in the pipe and the Reynolds number are 60 psi
V2 =
20/60 gal/s ⎛ 0.1337 ft V V ⎜ = → V2 = 2 Ac πD / 4 πD 2 / 4 ⎜⎝ 1 gal
Re =
(62.3 lbm/ft 3 )V 2 D ρV 2 D → Re = μ 6.556 × 10 − 4 lbm/ft ⋅ s
3
⎞ ⎟ ⎟ ⎠
⎞ V 22 ⎟= ⎟ 2(32.2 ft/s 2 ) + h L (1) ⎠ D 50 ft
(2)
2
1 (3)
Water main
60 psig
The friction factor can be determined from the Colebrook equation, ⎛ 0.00085 / D ⎛ ε / Dh 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ (4) → = −2.0 log⎜ + = −2.0 log⎜ + ⎜ ⎜ 3.7 3.7 Re f ⎟⎠ Re f ⎟⎠ f f ⎝ ⎝ The sum of the loss coefficients is K L = K L,entrance + 3K L,elbow + K L,gate valve + K L,angle valve = 0.5 + 3 × 1.1 + 0.2 + 5 = 9
∑
Then the total head loss becomes 2 V 22 ⎞V ⎛ 50 ft ⎞ ⎛ L (5) + → hL = ⎜ f + 9⎟ hL = ⎜ f KL ⎟ D ⎝ D ⎠ 2g ⎝ ⎠ 2(32.2 ft/s 2 ) These are 5 equations in the 5 unknowns of V2, hL, D, Re, and f, and solving them simultaneously using an equation solver such as EES gives hL = 135.5 ft, D = 0.0630 ft = 0.76 in, Re = 85,540, and f = 0.04263 V2 = 14.3 ft/s,
∑
Therefore, the diameter of the pipe must be at least 0.76 in (or roughly 3/4 in). Discussion The pipe diameter can also be determined approximately by using the Swamee and Jain relation. It would give D = 0.73 in, which is within 5% of the result obtained above.
8-96 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-126 Solution A water fountain is to be installed at a remote location by attaching a cast iron pipe directly to a water main. For a specified flow rate, the minimum diameter of the piping system is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed and the friction factor is constant for the entire pipe. 3 The pressure at the water main remains constant. 4 There are no dynamic pressure effects at the pipe-water main connection, and the pressure at the pipe entrance is 60 psia. 5 Elevation difference between the pipe and the fountain is negligible (z2 = z1) . 6 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The plastic pipes are considered to be smooth, and thus their roughness is ε = 0. The minor loss coefficient is KL = 0.5 for a sharp-edged entrance, KL = 1.1 for a 90° miter bend without vanes, KL = 0.2 for a fully open gate valve, and KL = 5 for an angle valve. Analysis We choose point 1 in the water main near the entrance where the pressure is 60 psig and the velocity in the pipe to be low. We also take point 1 as the reference level. We take point 2 at the exit of the water fountain where the pressure is the atmospheric pressure (P2 = Patm) and the velocity is the discharge velocity. The energy equation for a control volume between these two points is P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L 2g 2g ρg ρg
→
P1,gage
ρg
=α2
V 22 + hL 2g
2 ⎞ V2
⎛ L KL ⎟ where α2 = 1 and h L = h L ,total = h L,major + h L ,minor = ⎜ f + ⎝ D ⎠ 2g since the diameter of the piping system is constant. Then the energy equation becomes
∑
⎛ 144 lbf/ft 2 ⎞⎛ 32.2 lbm ⋅ ft/s 2 ⎜ ⎟⎜ 1 lbf (62.3 lbm/ft 3 )(32.2 ft/s 2 ) ⎜⎝ 1 psi ⎟⎠⎜⎝ The average velocity in the pipe and the Reynolds number are 60 psi
V2 =
20/60 gal/s ⎛ 0.1337 ft 3 V V ⎜ = → = V 2 Ac πD 2 / 4 πD 2 / 4 ⎜⎝ 1 gal
⎞ ⎟ ⎟ ⎠
⎞ V 22 ⎟= ⎟ 2(32.2 ft/s 2 ) + h L (1) ⎠
2 (2)
D
1
(62.3 lbm/ft 3 )V 2 D ρV 2 D → Re = Re = μ 6.556 × 10 − 4 lbm/ft ⋅ s
50 ft
(3)
The friction factor can be determined from the Colebrook equation,
Water main
60 psig
⎛ ε / Dh ⎛ 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ 0 + (4) ⎜ 3.7 ⎜ f f Re f ⎟⎠ Re f ⎟⎠ ⎝ ⎝ The sum of the loss coefficients is K L = K L,entrance + 3K L,elbow + K L,gate valve + K L,angle valve = 0.5 + 3 × 1.1 + 0.2 + 5 = 9 1
∑
Then the total head loss becomes 2 V 22 ⎞V ⎛ 50 ft ⎞ ⎛ L (5) + → hL = ⎜ f + 9⎟ hL = ⎜ f KL ⎟ D ⎝ D ⎠ 2g ⎝ ⎠ 2(32.2 ft/s 2 ) These are 5 equations in the 5 unknowns of V2, hL, D, Re, and f, and solving them simultaneously using an equation solver such as EES gives hL = 133.4 ft, D = 0.05549 ft = 0.67 in, Re = 97,170, and f = 0.0181 V2 = 18.4 ft/s,
∑
Therefore, the diameter of the pipe must be at least 0.67 in. Discussion The pipe diameter can also be determined approximately by using the Swamee and Jain relation. It would give D = 0.62 in, which is within 7% of the result obtained above.
8-97 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-127 Solution In a hydroelectric power plant, the flow rate of water, the available elevation head, and the combined turbine-generator efficiency are given. The electric power output of the plant is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed and the friction factor is constant for the entire pipe. 3 The minor losses are given to be negligible. 4 The water level in the reservoir remains constant. Properties The density and dynamic viscosity of water at 20°C are ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s. The roughness of cast iron pipes is ε = 0.00026 m. Analysis We take point 1 at the free surface of the reservoir, and point 2 and the reference level at the free surface of the water leaving the turbine site (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocities at both points are very low (V1 ≅ V2 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L → h turbine, e = z1 − h L 2g 2g ρg ρg The average velocity, Reynolds number, friction factor, and head loss in the pipe are 0.8 m 3 /s V V = = = 8.315 m/s V = Ac πD 2 / 4 π (0.35 m) 2 / 4 1 ρVD h (998 kg/m 3 )(8.513 m/s)(0.35 m) Re = = = 2.899 × 10 6 μ 70 m 1.002 × 10 − 3 kg/m ⋅ s Water
0.35 m which is greater than 4000. Therefore, the flow is turbulent. The relative roughness of the pipe is Turbine 200 m 0.00026 m −4 ε / Dh = = 7.43 × 10 0.35 m The friction factor can be determined from the Moody chart, but 2 to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎛ε / D ⎛ 7.43 × 10 −4 ⎞ 1 2.51 ⎞⎟ 1 2.51 h ⎟ = −2.0 log⎜ + → = −2.0 log⎜ + 6 ⎜ 3.7 ⎟ ⎜ ⎟ 3.7 f Re f f 2 . 899 10 f × ⎝ ⎠ ⎝ ⎠ It gives f = 0.0184. When the minor losses are negligible, the head loss in the pipe and the available turbine head are determined to be
hL = f
L V2 200 m (8.315 m/s) 2 = 0.0184 = 37.05 m D 2g 0.35 m 2(9.81 m/s 2 )
hturbine, e = z1 − h L = 70 - 37.05 = 32.95 m
Then the extracted power from water and the actual power output of the turbine become W turbine, e = m gh turbine, e = ρVgh turbine, e ⎛ 1 kN = (998 kg/m 3 )(0.8 m 3 /s)(9.81 m/s 2 )(32.95 m)⎜ ⎜ 1000 kg ⋅ m/s 2 ⎝ W = (0.84)(258 kW) = 217 kW
W turbine-gen = η turbine-gen
⎞⎛ 1 kW ⎞ ⎟⎜ = 258 kW ⎟⎝ 1 kN ⋅ m/s ⎟⎠ ⎠
turbine, e
Discussion Note that a perfect turbine-generator would generate 258 kW of electricity from this resource. The power generated by the actual unit is only 217 kW because of the inefficiencies of the turbine and the generator. Also note that more than half of the elevation head is lost in piping due to pipe friction.
8-98 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-128 Solution In a hydroelectric power plant, the flow rate of water, the available elevation head, and the combined turbine-generator efficiency are given. The percent increase in the electric power output of the plant is to be determined when the pipe diameter is tripled. Assumptions 1 The flow is steady and incompressible. 2 Entrance effects are negligible, and thus the flow is fully developed and friction factor is constant. 3 Minor losses are negligible. 4 Water level is constant. Properties The density and dynamic viscosity of water at 20°C are ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s. The roughness of cast iron pipes is ε = 0.00026 m. Analysis We take point 1 at the free surface of the reservoir, and point 2 and the reference level at the free surface of the water leaving the turbine site (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocities at both points are very low (V1 ≅ V2 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + h pump, u = 2 + α 2 2 + z 2 + h turbine, e + h L → hturbine, e = z1 − h L ρg 2g ρg 2g The average velocity, Reynolds number, friction factor, and head loss in the pipe for both cases (pipe diameter being 0.35 m and 1.05 m) are 0.8 m 3 /s V V = = = 8.315 m/s , V1 = Ac1 πD12 / 4 π (0.35 m) 2 / 4 1 0.8 m 3 /s V V = = = 0.9239 m/s V2 = 70 m Ac 2 πD 22 / 4 π (1.05 m) 2 / 4 Water
ρV1 D1 (998 kg/m 3 )(8.315 m/s)(0.35 m) Re 1 = = = 2.899 × 10 6 μ 1.002 × 10 −3 kg/m ⋅ s Re 2 =
1.05 m
Turbine 200 m
ρV 2 D1 (998 kg/m 3 )(0.9239 m/s)(1.05 m) = = 0.9662 × 10 6 −3 μ 1.002 × 10 kg/m ⋅ s
2 which are greater than 4000. Therefore, the flow is turbulent for both cases. The relative roughness of the pipe is 0.00026 m 0.00026 m and ε / D 2 = ε / D1 = = 7.43 × 10 − 4 = 2.476 × 10 − 4 0.35 m 1.05 m The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), ⎞ ⎞ ⎛ 7.43 × 10 −4 ⎛ 2.476 × 10 −4 2.51 1 2.51 ⎟ and ⎟ ⎜ 2 . 0 log = −2.0 log⎜ + = − + 6 6 ⎟ ⎟ ⎜ ⎜ 3 . 7 3 . 7 f f f1 2 . 899 10 f 0 . 9662 10 × × 2 1 ⎠ 2 ⎠ ⎝ ⎝ The friction factors are determined to be f1 = 0.01842 and f2 = 0.01520. When the minor losses are negligible, the head losses in the pipes and the head extracted by the turbine are determined to be 1
h L1 = f 1
L V12 200 m (8.315 m/s) 2 = 0.01842 = 37.09 m , hturbine,1 = z1 − h L1 = 70 - 37.09 = 32.91 m D1 2 g 0.35 m 2(9.81 m/s 2 )
L V 22 200 m (0.9239 m/s) 2 = 0.0152 = 0.13 m , hturbine,2 = z 2 − h L 2 = 70 - 0.13 = 69.87 m D2 2 g 1.05 m 2(9.81 m/s 2 ) The available or actual power output is proportional to the turbine head. Therefore, the increase in the power output when the diameter is doubled becomes hturbine,2 − h turbine,1 69.87 − 32.91 Increase in power output = = = 1.12 or 112% 32.91 hturbine,1 hL2 = f 2
Discussion Note that the power generation of the turbine more than doubles when the pipe diameter is tripled at the same flow rate and elevation.
8-99 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-129E Solution The drinking water needs of an office are met by siphoning water through a plastic hose inserted into a large water bottle. The time it takes to fill a glass when the bottle is first opened and when it is empty are to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed and the friction factor is constant for the entire pipe. 3 The on/off switch is fully open during filing. 4 The water level in the bottle remains nearly constant during filling. 5 The flow is turbulent (to be verified). 6 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The plastic pipes are considered to be smooth, and thus their roughness is ε = 0. The total minor loss coefficient is given to be 2.8. Analysis We take point 1 to be at the free surface of water in the bottle, and point 2 at the exit of the hose., which is also taken to be the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 = α 2
V 22 + hL 2g
2 V 22 6 ft ⎛ L ⎞V ⎛ ⎞ K L ⎟ 2 → hL = ⎜ f where α2 = 1 and h L = ⎜ f + + 2.8 ⎟ ⎠ 2g ⎝ 0.35 / 12 ft ⎠ 2(32.2 ft/s 2 ) ⎝ D since the diameter of the piping system is constant. Then the energy equation becomes V 22 (2) z1 = (1) + hL 2(32.2 ft/s 2 ) The average velocity in the pipe and the Reynolds number are
∑
V2 =
V V V ft 3 /s V = → = 2 Ac πD 2 / 4 π (0.35 / 12 ft ) 2 / 4
Re =
(62.3 lbm/ft 3 )V 2 (0.35 / 12 ft ) ρV 2 D → Re = μ 6.556 × 10 − 4 lbm/ft ⋅ s
(3)
(1)
6 ft
1
0.35
1 ft
(4) 3 ft
The friction factor can be determined from the Colebrook equation, 2 ⎛ε / D ⎛ 1 2.51 ⎞⎟ 1 2.51 ⎞⎟ h (5) = −2.0 log⎜ + → = −2.0 log⎜ 0 + ⎜ 3.7 ⎜ f f Re f ⎟⎠ Re f ⎟⎠ ⎝ ⎝ Finally, the filling time of the glass is V glass 0.00835 ft 3 (6) Δt = = V V ft 3 /s These are 6 equations in the 6 unknowns of V2, V , hL, Re, f, and Δt, and solving them simultaneously using an equation solver such as EES with the appropriate z1 value gives Case (a): The bottle is full and thus z1 = 3+1 = 4 ft: V2 = 5.185 ft/s, hL = 3.58 ft, V = 0.00346 ft3/s,
Re = 14,370, f = 0.02811, and Δt = 2.4 s
Case (b): The bottle is almost empty and thus z1 = 3 ft: hL = 2.69 ft, V = 0.00296 ft3/s, Re = 12,290, f = 0.02926, and Δt = 2.8 s V2 =4.436 ft/s,
Note that the flow is turbulent for both cases since Re > 4000. Discussion
The filling time of the glass increases as the water level in the bottle drops, as expected.
8-100 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes
8-130E
Solution In the previous problem, the effect of the hose diameter on the time required to fill a glass when the bottle is full is to be investigated by varying the pipe diameter from 0.2 to 2 in. in increments of 0.2 in. The EES Equations window is printed below, along with the tabulated and plotted results.
rho=62.3 mu=2.36/3600 nu=mu/rho g=32.2 z1=4 Volume=0.00835 D=Din/12 Ac=pi*D^2/4 L=6 KL=2.8 eps=0 rf=eps/D V=Vdot/Ac "Reynolds number" Re=V*D/nu 1/sqrt(f)=-2*log10(rf/3.7+2.51/(Re*sqrt(f))) HL=(f*L/D+KL)*(V^2/(2*g)) z1=V^2/(2*g)+HL Time=Volume/Vdot
10
8
Time, s
Analysis
6
4
2
0 0
0.4
0.8
1.2
1.6
2
D, in
D, in
Time, s
hL, ft
Re
0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0
9.66 1.75 0.68 0.36 0.22 0.15 0.11 0.08 0.06 0.05
3.76 3.54 3.40 3.30 3.24 3.20 3.16 3.13 3.11 3.10
6273 17309 29627 42401 55366 68418 81513 94628 107752 120880
Discussion The required time decreases considerably as the tube diameter increases. This is because the irreversible frictional head loss (major loss) in the tube decreases greatly as tube diameter increases. In addition, the minor loss is proportional to V 2. Thus, as tube diameter increases, V decreases, and even the minor losses decrease.
8-101 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-131E Solution The drinking water needs of an office are met by siphoning water through a plastic hose inserted into a large water bottle. The time it takes to fill a glass when the bottle is first opened is to be determined. Assumptions 1 The flow is steady and incompressible. 2 The entrance effects are negligible, and thus the flow is fully developed and the friction factor is constant for the entire pipe. 3 The on/off switch is fully open during filing. 4 The water level in the bottle remains constant during filling. 5 The flow is turbulent (to be verified). 6 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 70°F are ρ = 62.30 lbm/ft3 and μ = 2.360 lbm/ft⋅h = -4 6.556×10 lbm/ft⋅s. The plastic pipes are considered to be smooth, and thus their roughness is ε = 0. The total minor loss coefficient is given to be 2.8 during filling. Analysis We take point 1 to be at the free surface of water in the bottle, and point 2 at the exit of the hose, which is also taken to be the reference level (z2 = 0). Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 = α 2
V 22 + hL 2g
2 V 22 12 ft ⎛ L ⎞V ⎛ ⎞ K L ⎟ 2 → hL = ⎜ f where α2 = 1 and h L = ⎜ f + + 2.8 ⎟ ⎠ 2g ⎝ 0.35 / 12 ft ⎠ 2(32.2 ft/s 2 ) ⎝ D since the diameter of the piping system is constant. Then the energy equation becomes V 22 (2) z1 = (1) + hL 2(32.2 ft/s 2 ) The average velocity in the pipe and the Reynolds number are V V V ft 3 /s (3) V2 = V = → = 2 Ac πD 2 / 4 π (0.35 / 12 ft ) 2 / 4
∑
(62.3 lbm/ft 3 )V 2 (0.35 / 12 ft ) ρV 2 D Re = (4) → Re = μ 1.307 × 10 −3 lbm/ft ⋅ s The friction factor can be determined from the Colebrook equation, ⎛ε / D 2.51 ⎞⎟ h → = −2.0 log⎜ + ⎜ 3.7 Re f ⎟⎠ f ⎝ Finally, the filling time of the glass is V glass 0.00835 ft 3 Δt = = (6) V V ft 3 /s 1
⎛ 2.51 = −2.0 log⎜ 0 + ⎜ Re f f ⎝
1
⎞ ⎟ ⎟ ⎠
(1)
12 ft 0.35 in
1
1 ft
3 ft
(5)
2
These are 6 equations in the 6 unknowns of V2, V , hL, Re, f, and Δt, and solving them simultaneously using an equation solver such as EES with the appropriate z1 value gives Case (a): The bottle is full and thus z1 = 3+1 = 4 ft: hL = 3.75 ft, V = 0.002667 ft3/s, Re = 11,060, f = 0.03007, and Δt = 3.1 s V2 =3.99 ft/s, Case (b): The bottle is almost empty and thus z1 = 3 ft: hL = 2.82 ft, V = 0.002272 ft3/s, V2 =3.40 ft/s,
Re = 9426, f =0.03137, and Δt = 3.7 s
Note that the flow is turbulent for both cases since Re > 4000. Discussion The filling times in Prob. 8-129E were 2.4 s and 2.8 s, respectively. Therefore, doubling the tube length increases the filling time by 0.7 s when the bottle is full, and by 0.9 s when it is empty.
8-102 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 8 Flow in Pipes 8-132 Solution A water pipe has an abrupt expansion from diameter D1 to D2. It is to be shown that the loss coefficient is 2 K L = (1 − D1 / D22 ) 2 , and KL and P2 are to be calculated. Assumptions 1 The flow is steady and incompressible. 2 The pressure is uniform at the cross-section where expansion occurs, and is equal to the upstream pressure P1. 3 The flow section is horizontal (or the elevation difference across the expansion section is negligible). 4 The flow is turbulent, and the effects of kinetic energy and momentum-flux correction factors are negligible, β ≈ 1 and α ≈ 1.
We take the density of water to be ρ = 1000 kg/m3.
Properties
Analysis We designate the cross-section where expansion occurs by x. We choose cross-section 1 in the smaller diameter pipe shortly before x, and section 2 in the larger diameter pipe shortly after x. We take the region occupied by the fluid between cross-sections 1 and 2 as the control volume, with an inlet at 1 and exit at 2. The velocity, pressure, and cross-sectional area are V1, P1, and A1 at cross-section 1, and V2, P2, and A2 at cross-section 2. We assume the pressure along the cross-section x to be P1 so that Px = P1. Then the continuity, momentum, and energy equations applied to the control volume become A (1) Continuity: m 1 = m 2 → ρV1 A1 = ρV 2 A2 → V 2 = 1 V1 (1) A2 G F= βm V − β m V → P1 A1 + P1 ( Ax − A1 ) − P2 A2 = m (V 2 − V1 ) (2) Momentum:
∑
∑ out
∑ in
P1 A1 + P1 ( Ax − A1 ) = P1 Ax = P1 A2
But
m (V 2 − V1 ) = ρA2V 2 (V 2 − V1 ) = ρA2 P1 A2 − P2 A2 = ρA2
Therefore, (3) Energy:
A1 A2
⎞ ⎞ A1 ⎛ A1 A ⎛A V1 ⎜⎜ V1 − V1 ⎟⎟ = ρA2 1 ⎜⎜ 1 − 1⎟⎟V12 A2 ⎝ A2 A2 ⎝ A2 ⎠ ⎠
⎛ A1 ⎞ ⎜⎜ − 1⎟⎟V12 → ⎝ A2 ⎠
P1 − P2
ρ
=
A1 A2
⎛ A1 ⎞ ⎜⎜ − 1⎟⎟V12 ⎝ A2 ⎠
(2)
P1 V2 P V2 P − P2 V12 − V 22 + α 1 1 + z1 + h pump,u = 2 + α 2 2 + z 2 + hturbine, e + h L → h L = 1 + (3) ρg 2g ρg 2g ρg 2g
Substituting Eqs. (1) and (2) and h L = K L
V12 into Eq. (3) gives 2g
⎛ A1 ⎞ V 2 V 2 − ( A12 / A22 )V12 ⎞ ⎛ A2 ⎞ 2A ⎛ A ⎜⎜ − 1⎟⎟ 1 + 1 → K L = 1 ⎜⎜ 1 − 1⎟⎟ + ⎜1 − 12 ⎟ ⎜ ⎟ 2g A2 ⎝ A2 ⎝ A2 ⎠ g ⎠ ⎝ A2 ⎠ Simplifying and substituting A = πD2/4 gives the desired relation and its value, KL
V12 A = 1 2 g A2
2
2 ⎛ ⎛ πD 2 / 4 ⎞ ⎛ D2 ⎞ ⎛ A ⎞ ⎟ = ⎜ 1 − 1 ⎟ = ⎜ 1 − (0.15 m) K L = ⎜ 1 − 1 ⎟ = ⎜⎜ 1 − 12 2 2 ⎜ ⎟ ⎜ D ⎟ ⎜ A1 ⎟⎠ ⎝ (0.20 m) 2 ⎠ ⎝ πD 2 / 4 ⎠ ⎝ ⎝
Also,
hL = K L
2
2
2
⎞ ⎟ = 0.1914 ⎟ ⎠
V12 (10 m/s ) 2 = (0.1914) = 0.9756 m 2g 2(9.81 m/s 2 )
A D2 (0.15 m) 2 V 2 = 1 V1 = 12 V1 = (10 m/s) = 5.625 m/s A2 D2 (0.20 m) 2 Solving for P2 from Eq. (3) and substituting, P2 = P1 + ρ{(V1 2 − V 2 2 ) / 2 − gh L }
1 10 m/s
15 cm
2 20 cm
⎧ (10 m/s ) 2 − (5.625 m/s ) 2 ⎫⎛ 1 kPa ⋅ m 2 ⎞ ⎟ = 145 kPa = (120 kPa) + (1000 kg/m 3 )⎨ − (9.81 m/s 2 )(0.9756 m)⎬⎜⎜ 2 ⎟ 2 ⎩ ⎭⎝ 1000 kg ⋅ m/s ⎠ Note that the pressure increases by 25 kPa after the expansion due to the conversion of dynamic pressure to static pressure when the velocity is decreased. Also, KL ≅ 1 (actually, KL = α) when D2 >> D1 (discharging into a reservoir). Discussion
At a discharge into a large reservoir, all the kinetic energy is wasted as heat.
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Chapter 8 Flow in Pipes 8-133 Solution A swimming pool is initially filled with water. A pipe with a well-rounded entrance at the bottom drains the pool to the atmosphere. The initial rate of discharge from the pool and the time required to empty the pool completely are to be determined. Assumptions 1 The flow is uniform and incompressible. 2 The draining pipe is horizontal. 3 The entrance effects are negligible, and thus the flow is fully developed. 4 The friction factor remains constant (in reality, it changes since the flow velocity and thus the Reynolds number changes during flow). 5 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 20°C are ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s. The friction factor of the pipe is given to be 0.022. Plastic pipes are considered to be smooth, and their surface roughness is ε = 0. Analysis We take point 1 at the free surface of the pool, and point 2 and the reference level at the exit of the pipe (z2 = 0), and take the positive direction of z to be upwards. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 = α 2
V 22 + hL 2g
where L V2 D 2g
hL = f
10 m
1
since the minor losses are negligible. Substituting and solving for V2 gives z1 = α 2
V 22 ⎛ L ⎞ V 22 +⎜ f ⎟ 2g ⎝ D ⎠ 2g
→
V2 =
2 gz1
α 2 + fL / D
2m
Swimming pool
Noting that α2 = 1 and initially z1 = 2 m, the initial velocity and flow rate are determined to be V 2 ,i
2 gz1 2(9.81 m/s 2 )(2 m) = = = 1.425 m/s 1 + fL / D 1 + 0.022(25 m)/(0.03 m)
2 25 m
3 cm
Vinitial = V 2,i Ac = V 2,i (πD 2 / 4) = (1.425 m/s)[π (0.03 m) 2 /4] = 1.01× 10 −3 m 3 /s = 1.01 L/s
The average discharge velocity at any given time, in general, can be expressed as V2 =
2 gz 1 + fL / D
where z is the water height relative to the center of the orifice at that time. We denote the diameter of the pipe by D, and the diameter of the pool by Do. The flow rate of water from the pool can be obtained by multiplying the discharge velocity by the pipe cross-sectional area,
πD V = AcV 2 = 4
2
2 gz 1 + fL / D
Then the amount of water that flows through the pipe during a differential time interval dt is dV = Vdt =
πD 2 4
2 gz dt 1 + fL / D
(1)
which, from conservation of mass, must be equal to the decrease in the volume of water in the pool, dV = Ac ,tan k (− dz ) = −
πD02 4
dz
(2)
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Chapter 8 Flow in Pipes where dz is the change in the water level in the pool during dt. (Note that dz is a negative quantity since the positive direction of z is upwards. Therefore, we used –dz to get a positive quantity for the amount of water discharged). Setting Eqs. (1) and (2) equal to each other and rearranging, D2 πD 2 2 gz dt = − 0 dz → dt = − 02 1 + fL / D 4 D
πD 2 4
D2 1 + fL / D dz = − 02 2 gz D
1 + fL / D − 12 z dz 2g
The last relation can be integrated easily since the variables are separated. Letting tf be the discharge time and integrating it from t = 0 when z = z1 to t = tf when z = 0 (completely drained pool) gives
∫
tf
dt = −
t =0
D 02 D
2
1 + fL / D 2g
∫
0
z
−1 / 2
z = z1
dz → t f = -
D 02 D
2
1
1 + fL / D z 2 1 2g 2
0
=
2 D02
z1
D
2
1 + fL / D 12 z1 2g
Simplifying and substituting the values given, the draining time is determined to be tf =
D02 D
2
2 z1 (1 + fL / D ) (10 m) 2 = g (0.03 m) 2
2(2 m)[1 + (0.022)(25 m)/(0.03 m)] 9.81 m/s 2
= 312,000 s = 86.7 h
Checking: For plastic pipes, the surface roughness and thus the roughness factor is zero. The Reynolds number at the beginning of draining process is
ρV 2 D (998 kg/m 3 )(1.425 m/s)(0.03 m) = = 42,580 μ 1.002 × 10 −3 kg/m ⋅ s which is greater than 4000. The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), Re =
⎛ε / D ⎛ 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ 0 + ⎜ 3.7 Re f ⎟ ⎜ f f 42,570 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.022. Therefore, the given value of 0.022 is accurate. 1
Discussion It can be shown by setting L = 0 that the draining time without the pipe is only about 18 h. Therefore, the pipe in this case increases the draining time by about a factor of 5.
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Chapter 8 Flow in Pipes
8-134
Solution In the previous problem, the effect of the discharge pipe diameter on the time required to empty the pool completely is to be investigated by varying the pipe diameter from 1 cm to 10 cm in increments of 1 cm. Analysis
The EES Equations window is printed below, along with the tabulated and plotted results.
rho=998 mu=0.001002 g=9.81 Dtank= 10 Ac=pi*D^2/4 L=25 f=0.022 z1=2 V=(2*g*z1/(1+f*L/D))^0.5 Vdot=V*Ac Time=(Dtank/D)^2*(2*z1*(1+f*L/D)/g)^0.5/3600
D, m
Time, h
Vinitial, m/s
Re
0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10
1327.4 236.7 86.7 42.6 24.6 15.7 10.8 7.8 5.8 4.5
0.84 1.17 1.42 1.63 1.81 1.96 2.10 2.23 2.35 2.46
8337 23374 42569 64982 90055 117406 146750 177866 210572 244721
1400 1200
Time, h
1000 800 600 400 200 0 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09
0.1
D, m Discussion
The required drain time decreases quite rapidly as pipe diameter is increased.
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Chapter 8 Flow in Pipes 8-135
Solution A swimming pool is initially filled with water. A pipe with a sharp-edged entrance at the bottom drains the pool to the atmosphere. The initial rate of discharge from the pool and the time required to empty the pool completely are to be determined. Assumptions 1 The flow is uniform and incompressible. 2 The draining pipe is horizontal. 3 The flow is turbulent so that the tabulated value of the loss coefficient can be used. 4 The friction factor remains constant (in reality, it changes since the flow velocity and thus the Reynolds number changes during flow). 5 The effect of the kinetic energy correction factor is negligible, α = 1. Properties The density and dynamic viscosity of water at 20°C are ρ = 998 kg/m3 and μ = 1.002×10-3 kg/m⋅s. The loss coefficient for the sharp-edged entrance is KL = 0.5. Plastic pipes are considered to be smooth, and their surface roughness is ε = 0. Analysis We take point 1 at the free surface of the pool, and point 2 and the reference level at the exit of the pipe (z2 = 0), and take the positive direction of z to be upwards. Noting that the fluid at both points is open to the atmosphere (and thus P1 = P2 = Patm) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the energy equation for a control volume between these two points (in terms of heads) simplifies to P1 V2 P V2 + α 1 1 + z1 + hpump, u = 2 + α 2 2 + z 2 + h turbine, e + h L ρg 2g ρg 2g
→
z1 = α 2
V 22 + hL 2g
where α2 = 1 and ⎛ L h L = h L ,total = h L ,major + h L ,minor = ⎜ f + ⎝ D
∑K
L
2 2 ⎞V ⎛ L ⎞V =⎜ f + KL ⎟ 2 ⎟ ⎠ 2g ⎝ D ⎠ 2g
since the diameter of the piping system is constant. Substituting and solving for V2 gives 2 2 gz1 V 22 ⎛ L ⎞V +⎜ f + KL ⎟ 2 → V2 = 2g ⎝ D 1 + fL / D + K L ⎠ 2g Noting that initially z1 = 2 m, the initial velocity and flow rate are determined to be
z1 =
V 2 ,i =
2 gz1 2(9.81 m/s 2 )(2 m) = = 1.407 m/s 1 + fL / D + K L 1 + 0.022(25 m)/(0.03 m) + 0.5
Vinitial = V 2,i Ac = V 2,i (πD 2 / 4) = (1.407 m/s)[π (0.03 m) 2 /4] = 9.94 × 10 −4 m 3 /s = 0.994 L/s
The average discharge velocity at any given time, in general, can be expressed as V2 =
2 gz 1 + fL / D + K L
where z is the water height relative to the center of the orifice at that time. We denote the diameter of the pipe by D, and the diameter of the pool by Do. The flow rate of water from the pool can be obtained by multiplying the discharge velocity by the pipe cross-sectional area,
πD V = AcV 2 = 4
2
2 gz 1 + fL / D + K L
10 m
1
Then the amount of water that flows through the pipe during a differential time interval dt is
πD dV = Vdt = 4
2
2m
2 gz dt 1 + fL / D + K L
(1)
2
which, from conservation of mass, must be equal to the decrease in the volume of water in the pool, dV = Ac ,tan k (− dz ) = −
πD02 4
dz
Swimming pool
25 m
3 cm
(2)
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Chapter 8 Flow in Pipes where dz is the change in the water level in the pool during dt. (Note that dz is a negative quantity since the positive direction of z is upwards. Therefore, we used –dz to get a positive quantity for the amount of water discharged). Setting Eqs. (1) and (2) equal to each other and rearranging, D 2 1 + fL / D + K L D 2 1 + fL / D + K L − 12 πD 2 2 gz dt = − 0 dz → dt = − 02 dz = − 02 z dz The last relation 4 1 + fL / D + K L 4 2 gz 2g D D can be integrated easily since the variables are separated. Letting tf be the discharge time and integrating it from t = 0 when z = z1 to t = tf when z = 0 (completely drained pool) gives
πD 2
∫
tf
dt = −
t =0
D02 D2
1 + fL / D + K L 2g
∫
0
z
−1 / 2
z = z1
dz → t f = -
D02 D2
1
1 + fL / D + K L z 2 1 2g 2
0
= z1
2 D02 D2
1 + fL / D + K L 12 z1 Simplifying 2g
and substituting the values given, the draining time is determined to be 2 z1 (1 + fL / D + K L ) (10 m) 2 2(2 m)[1 + (0.022)(25 m)/(0.03 m) + 0.5] = = 316,000 s = 87.8 h This is a 2 2 g (0.03 m) 9.81 m/s 2 D change of (87.8-86.7)/86.7 = 0.013 or 1.3%. Therefore, the minor loss in this case is truly minor. tf =
D02
Checking: For plastic pipes, the surface roughness and thus the roughness factor is zero. The Reynolds number at the beginning of draining process is
ρV 2 D (998 kg/m 3 )(1.407 m/s)(0.03 m) = = 42,030 μ 1.002 × 10 −3 kg/m ⋅ s which is greater than 4000. The friction factor can be determined from the Moody chart, but to avoid the reading error, we determine it from the Colebrook equation using an equation solver (or an iterative scheme), Re =
⎛ε / D ⎛ 2.51 ⎞⎟ 1 2.51 ⎞⎟ = −2.0 log⎜ + → = −2.0 log⎜ 0 + ⎜ 3.7 Re f ⎟ ⎜ f f 42,030 f ⎟⎠ ⎝ ⎠ ⎝ It gives f = 0.022. Therefore, the given value of 0.022 is accurate. 1
Discussion It can be shown by setting L = 0 that the draining time without the pipe is only about 24 h. Therefore, the pipe in this case increases the draining time more than 3 folds.
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Chapter 8 Flow in Pipes 8-136 Solution A system that concsists of two interconnected cylindrical tanks is used to determine the discharge coefficient of a short 5-mm diameter orifice. For given initial fluid heights and discharge time, the discharge coefficient of the orifice is to be determined. Assumptions 1 The fluid is incompressible. 2 The entire systems, including the connecting flow section, is horizontal. 3 The discharge coefficient remains constant (in reality, it may change since the flow velocity and thus the Reynolds number changes during flow). 4 Losses other than the ones associated with flow through the orifice are negligible. 5 The effect of the kinetic energy correction factor is negligible, α = 1. Analysis We take point 1 at the free surface of water in Tank 1, and point 0 at the exit of the orifice. We take the centerline of the orifice as the reference level (z1 = h1 and z0 = 0). Noting that the fluid at point 1is open to the atmosphere (and thus P1 = Patm and P0 = Patm + ρgh2) and that the fluid velocity at the free surface is very low (V1 ≅ 0), the Bernoulli equation between these two points gives P1 V12 P V2 P P + ρgh2 V02 + + z1 = 0 + 0 + z0 → atm + h1 = atm + 2g ρg 2 g ρg 2 g ρg ρg
h1
h
h2 Tank 1
Tank 2 Orifice
→ V0 = 2 g (h1 − h2 ) = 2 gh
where h = h1 − h2 is the vertical distance between the water levels in the two tanks at any time t. Note that h1, h2, h, and V0 are all variable (h1 decreases while h2 and h increase during discharge. Noting that the fluid is a liquid (ρ = constant) and keeping the conservation of mass in mind and the definition of the discharge coefficient Cd, the flow rate through the orifice can be expressed as dh dh A V = C d Vo Ao = − A1 1 = A2 2 ⎯→ dh2 = − 1 dh1 dt dt A2 Also,
h = h1 − h2
⎯→ dh = dh1 − dh2 ⎯→ dh1 = dh2 + dh (Note that dh < 0, dh1 < 0, and dh2>0) dh1 =
Combining the two equations above,
dh 1 + A1 A2
dh 1 dh V = C d Vo Ao = − A1 1 ⎯→ C d Ao 2 gh = − A1 dt 1 + A1 / A2 dt
Then,
A1 A2 1 dh A1 + A2 C d Ao 2 g h
− dt =
which can be rearranged as
t
∫
− dt =
Integrating
0
t=−
Performing the integration
Cd =
Solving for Cd
A1 A2 1 A1 + A2 C d Ao 2 g
A1 A2 2 A1 + A2 C d Ao 2 g
[
h
∫
h1
dh h
h − h1
]
2( h1 − h ) ( A0 / A2 + A0 / A1 )t 2 g
Fluid flow stops when the liquid levels in the two tanks become equal (and thus h = 0). Substituign the given values, the discharge coefficient is determined to be 2
2
2 2 A0 A0 ⎛ D0 ⎞ ⎛ D0 ⎞ 0.5 cm ⎞ ⎛ 0.5 cm ⎞ ⎟⎟ + ⎜⎜ ⎟⎟ = ⎛⎜ + = ⎜⎜ ⎟ +⎜ ⎟ = 0.002014 , A2 A1 ⎝ D 2 ⎠ ⎝ D1 ⎠ ⎝ 30 cm ⎠ ⎝ 12 cm ⎠
Cd =
Discussion
2 0.5 m (0.002014)(170 s) 2 × 9.81 m/s 2
= 0.933
We could add the minor losses at the pipe inlet and outlet without much extra effort.
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Chapter 8 Flow in Pipes 8-137 Solution A highly viscous liquid discharges from a large container through a small diameter tube in laminar flow. A relation is to be obtained for the variation of fluid depth in the tank with time.
Discharge tube
h
d
D
L
Assumptions 1 The fluid is incompressible. 2 The discharge tube is horizontal, and the flow is laminar. 3 Entrance effects and the velocity heads are negligible. Analysis We take point 1 at the free surface of water in the tank, and point 2 at the exit of the pipe. We take the centerline of the pipe as the reference level (z1 = h and z2 = 0). Noting that the fluid at both points 1 and 2 are open to the atmosphere (and thus P1 = P2 = Patm ) and that the fluid velocity at the free surface is very low (V1 ≅ 0) and the velocity heads are negligible, the energy equation for a control volume between these two points gives P1 V2 P V2 + α1 1 + z1 = 2 + α 2 2 + z2 + hL 2g 2g ρg ρg
→
Patm P + h = atm + hL ρg ρg
→
hL = h
(1)
where h is the liquid height in the tank at any time t. The total head loss through the pipe consists of major losses in the pipe since the minor losses are negligible. Also, the entrance effects are negligible and thus the friction factor for the entire tube is constant at the fully developed value. Noting that f = Re/ 64 for fully developed laminar flow in a circular pipe of diameter d, the head loss can be expressed as L V 2 64 L V 2 64 L V 2 64νL V (2) = = = 2 hL = f d 2 g Re d 2 g Vd / ν d 2 g d 2g V V = . Substituting into (2), The average velocity can be expressed in terms of the flow rate as V = Ac πd 2 / 4 64νL 1 ⎛ V ⎞ 64νL 4V 128νLV ⎜ ⎟= hL = 2 = (3) d 2 g ⎜⎝ πd 2 / 4 ⎟⎠ d 2 2 gπd 2 gπd 4 128νLV h= Combining Eqs. (1) and (3): (4) gπd 4 Noting that the liquid height h in the tank decreases during flow, the flow rate can also be expressed in terms of the rate of change of liquid height in the tank as dh πD 2 dh V = − Atank =− (5) 4 dt dt 128νL πD 2 dh 32νLD 2 dh =− (6) Substituting Eq. (5) into (4): h = − 4 4 dt gπd gd 4 dt 32νLD 2 dh h gd 4 Integrating from t = 0 (at which h = H) to t = t (at which h = h) gives 32νLD 2 t= ln( H / h ) gd 4 which is the desired relation for the variation of fluid depth h in the tank with time t.
To separate variables, it can be rearranged as
Discussion
dt = −
If the entrance effects and the outlet kinetic energy were included in the analysis, the time would be slower.
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Chapter 8 Flow in Pipes 8-138 Solution of data.
Using the setup described in the previous problem, the viscosity of an oil is to be determined for a given set
Discharge tube
h
d
D
L
Assumptions 1 The oil is incompressible. 2 The discharge tube is horizontal, and the flow is laminar. 3 Entrance effects and the inlet and the exit velocity heads are negligible.
The variation of fluid depth h in the tank with time t was determined in the previous problem to be
Analysis
t=
32νLD 2 gd 4
ln( H / h )
Solving for ν and substituting the given values, the kinematic viscosity of the oil is determined to be
ν=
gd 4 32 LD 2 ln( H / h)
t=
(9.81 m/s 2 )(0.006 m) 4 32(0.65 m)(0.63 m) 2 ln(0.4/0.36)
(2842 s) = 4.15 × 10 - 5 m 2 /s
Discussion Note that the entrance effects are not considered, and the velocity heads are disregarded. Also, the value of the viscosity strongly depends on temperature, and thus the oil temperature should be maintained constant during the test.
Design and Essay Problems
8-139 to 8-142 Solution Students’ essays and designs should be unique and will differ from each other.
KJ
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Chapter 9 Differential Analysis of Fluid Flow
Solutions Manual for Fluid Mechanics: Fundamentals and Applications by Çengel & Cimbala
CHAPTER 9 DIFFERENTIAL ANALYSIS OF FLUID FLOW
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Chapter 9 Differential Analysis of Fluid Flow General and Mathematical Problems
9-1C Solution
We are to explain the fundamental differences between a flow domain and a control volume.
Analysis A control volume is used in an integral, control volume solution. It is a volume over which all mass flow rates, forces, etc. are specified over the entire control surface of the control volume. In a control volume analysis we do not know or care about details inside the control volume. Rather, we solve for gross features of the flow such as net force acting on a body. A flow domain, on the other hand, is also a volume, but is used in a differential analysis. Differential equations of motion are solved everywhere inside the flow domain, and we are interested in all the details inside the flow domain. Discussion Note that we also need to specify what is happening at the boundaries of a flow domain – these are called boundary conditions.
9-2C Solution
We are to explain what we mean by coupled differential equations.
Analysis A set of coupled differential equations simply means that the equations are dependent on each other and must be solved together rather than separately. For example, the equations of motion for fluid flow involve velocity variables in both the conservation of mass equation and the momentum equation. To solve for these variables, we must solve the coupled set of differential equations together. Discussion
In some very simple fluid flow problems, the equations become uncoupled, and are easier to solve.
9-3C Solution We are to discuss the number of unknowns and the equations needed to solve for those unknowns for a three-dimensional, unsteady, incompressible flow field. Analysis equations: − −
There are four unknowns (velocity components u, v, w, and pressure P) and thus we need to solve four
one from conservation of mass which is a scalar equation three from Newton’s second law which is a vector equation
Discussion
These equations are also coupled in general.
9-4C Solution We are to discuss the number of unknowns and the equations needed to solve for those unknowns for a three-dimensional, unsteady, compressible flow field with significant variations in both temperature and density. Analysis equations: − − − −
There are six unknowns (velocity components u, v, w, ρ, T, and P) and thus we need to solve six
one from conservation of mass which is a scalar equation three from Newton’s second law which is a vector equation one from the energy equation which is a scalar equation one from an equation of state (e.g. ideal gas law) which is a scalar equation
Discussion
These equations are also coupled in general.
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Chapter 9 Differential Analysis of Fluid Flow 9-5C Solution
We are to express the divergence theorem in words. G G Analysis For vector G , the volume integral of the divergence of G over volume V is equal to the surface G integral of the normal component of G taken over the surface A that encloses the volume. Discussion
The divergence theorem is also called Gauss’s theorem.
9-6 Solution
We are to transform a position from Cartesian to cylindrical coordinates.
Analysis
We use the coordinate transformations provided in this chapter,
r = x2 + y 2 =
( 4 m ) + (3 m ) 2
2
=5m
(1)
y ⎛3m⎞ o = tan −1 ⎜ ⎟ = 36.87 = 0.6435 radians x ⎝4 m⎠
(2)
and
θ = tan −1
Coordinate z remains unchanged. Thus, G Position in cylindrical coordinates: x = ( r ,θ ,z ) = ( 5 m, 0.6435 radians, - 4 m )
(3)
Discussion
Notice that the units of θ are radians since angles are dimensionless.
9-7 Solution exact value.
We are to calculate a truncated Taylor series expansion for a given function and compare our result with the
Analysis
The algebra here is simple since d(ex)/dx = ex. The Taylor series expansion is 1 x0 2 1 x0 3 e dx + e dx + ... 2 3× 2
(1)
Truncated Taylor series expansion: 1 1 f (−0.1) ≈ 1 + 1× (−0.1) + × 1× (−0.1) 2 + × 1× (−0.1) 3 = 0.9048333... 2 6
(2)
Taylor series expansion:
f ( x0 + dx) = e x0 + e x0 dx +
We plug x0 = 0 and dx = –0.1 into Eq. 1,
We compare Eq. 2 with the exact value, Exact value:
f (−0.1) = e −0.1 = 0.904837418...
(3)
Comparing Eqs. 2 and 3 we see that our approximation is good to four or five significant digits. Discussion The smaller the value of dx, the better the approximation. You can easily convince yourself of this by trying dx = 0.01 instead.
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Chapter 9 Differential Analysis of Fluid Flow 9-8 Solution Analysis
We are to calculate the divergence of a given vector. G G G ∂ G ∂ G ∂ G The divergence of G is the dot product of the del operator ∇ = i + j + k with G , which gives ∂x ∂y ∂z
G⎞ G G ⎛ ∂ G ∂ G ∂ G⎞ ⎛ G 1 G ∇ ⋅G = ⎜ i + j + k ⎟ ⋅ ⎜ 2 xzi − x 2 j + z 2 k ⎟ = 2 z + 0 − 2 z = 0 ∂y ∂z ⎠ ⎝ 2 ⎠ ⎝ ∂x G It turns out that for this special case, the divergence of G is zero. G Discussion If G were a velocity vector, this would mean that the flow field is incompressible.
G Divergence of G :
9-9 Solution We are to perform both integrals of the divergence theorem for a given vector and volume, and verify that they are equal. Analysis
We do the volume integral first:
Volume integral:
G G ∂G y ∂Gz ⎞ x =1 y =1 z =1 ⎛ ∂G x =1 y =1 z =1 dzdydx = ∫ ∫ ∫ ( 4 z − 2 y + y )dzdydx ∇ ⋅ GdV = ∫ ∫ ∫ ⎜ x + + ⎟ x =0 y =0 z =0 x =0 y =0 z =0 V ∂y ∂z ⎠ ⎝ ∂x
∫
(1)
The term in parentheses in Eq. 1 reduces to (4z – y), and we integrate this over z first, G G x =1 y =1 x =1 y =1 z =1 2 ∫V ∇ ⋅ GdV = ∫x =0 ∫y =0 ⎡⎣2 z − yz ⎤⎦ z =0 dydx = ∫x =0 ∫y =0 ( 2 − y ) dydx Then we integrate over y and then over x, y =1 G G x =1 ⎡ x =1 3 y2 ⎤ 3 ∫V ∇ ⋅ GdV = ∫x =0 ⎢⎣2 y − 2 ⎥⎦ dx = ∫x =0 2 dx = 2 y =0
Volume integral:
(2)
G Next we calculate the surface integral of the divergence theorem. There are six faces of the cube, and unit vector n points G outward from each face. So, we split the area integral into six parts and sum them. E.g., the right-most face has n = (1,0,0), G G G G G so G ⋅ n = 4xz on this face. The bottom face has n = (0,–1,0), so G ⋅ n = y2 on this face. The surface integral is then
v∫ Surface integral:
G G y =1 z =1 y =1 z =1 z =1 x =1 − y 2 dxdz ⎤ G ⋅ ndA = ⎡ ∫ ∫ ( 4 xz ) dzdy ⎤ + ⎡ ∫ ∫ ( −4 xz ) dzdy ⎤ + ⎡ ∫ ∫ A y =0 z =0 ⎥ x =1 ⎣⎢ y = 0 z = 0 ⎥ x = 0 ⎣⎢ z = 0 x = 0 ⎣⎢ ⎦ ⎦ ⎦⎥ y =1
(
Right face
Left face
( )
)
Top face
+ ⎡∫ ∫ y 2 dxdz ⎤ + ⎡ ∫ ∫ ( yz ) dydx ⎤ + ⎡ ∫ ∫ ( − yz ) dydx ⎤ x =0 y =0 x =0 y =0 ⎣⎢ z = 0 x = 0 ⎦⎥ y = 0 ⎣⎢ ⎦⎥
⎣⎢ ⎦⎥ z
z =1 =0
z =1
x =1
x =1
y =1
Front face
Bottom face
x =1
y =1
(3)
Back face
The three integrals on the far right of Eq. 3 are obviously zero. The other three integrals can be obtained carefully, y =1 2 G G y =1 z =1 x =1 ⎡ y ⎤ y =1 z =1 x =1 ⎛ 1 ⎞ z =1 x =1 2 v∫ A G ⋅ ndA = ∫y =0 ⎡⎣ 2 z ⎤⎦ z =0 dy + ∫z =0 [ − x ]x =0 dz + ∫x =0 ⎢ 2 ⎥ dx = ∫y = 0 ( 2 ) dy + ∫z =0 ( −1) dz + ∫x =0 ⎜⎝ 2 ⎟⎠ dx ⎣ ⎦ y =0
(4)
The last three integrals of Eq. 4 are trivial. The final result is
Surface integral:
v∫
A
G G 1 3 G ⋅ ndA = 2 − 1 + = 2 2
(5)
Since Eq. 2 and Eq. 5 are equal, the divergence theorem works for this case. Discussion The integration is simple in this example since each face is flat and normal to an axis. In the general case in which the surface is curved, integration is much more difficult, but the divergence theorem always works.
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Chapter 9 Differential Analysis of Fluid Flow 9-10 Solution
We are to expand a dot product in Cartesian coordinates and verify it.
G G G ∂ G ∂ G ∂ G G G In Cartesian coordinates the del operator is ∇ = i + j + k and we let G = Gx i + Gy j + Gz k . The ∂x ∂y ∂z left hand side of the equation is thus
Analysis
(
( )
Left hand side:
)
G G ∂ ( fGx ) ∂ fGy ∂ ( fGz ) ∇ ⋅ fG = + + ∂x ∂y ∂z ∂G y ∂Gx ∂Gz ∂f ∂f ∂f = Gx +f + Gy +f + Gz +f ∂x ∂x ∂y ∂y ∂z ∂z
(1)
The right hand side of the equation is G G G G G ⋅ ∇f + f ∇ ⋅ G
Right hand side:
G ⎛ ∂f G ∂f G ∂f G ⎞ G G = Gx i + Gy j + Gz k ⋅ ⎜ i + j+ k ⎟+ ∂y ∂z ⎠ ⎝ ∂x ∂G y ∂Gx ∂f ∂f ∂f = Gx + Gy + Gz +f +f + ∂x ∂y ∂z ∂x ∂y
(
)
∂G y ∂Gz ⎞ ⎛ ∂G + f⎜ x + ⎟ ∂y ∂z ⎠ ⎝ ∂x ∂Gz f ∂z
(2)
Equations 1 and 2 are the same, and the given equation is verified. Discussion The product rule given in this problem was used in this chapter in the derivation of the alternative form of the continuity equation.
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Chapter 9 Differential Analysis of Fluid Flow 9-11 Solution
We are to expand the given equation in Cartesian coordinates and verify it. G G G ∂ G ∂ G ∂ G G G Analysis In Cartesian coordinates the del operator is ∇ = i + j + k and we let F = Fx i + Fy j + Fz k and ∂x ∂y ∂z G G G G G = Gx i + Gy j + Gz k . The left hand side of the equation is thus Fx Gz ⎤ ⎥ Fy G y Fy Gz ⎥ Fz Gy Fz Gz ⎥⎦ ⎡∂ ⎤G ∂ ∂ Fy Gx + ( Fz Gx ) ⎥ i = ⎢ ( Fx Gx ) + ∂y ∂z ⎣ ∂x ⎦ ⎡∂ ⎤G ∂ ∂ Fx G y + Fy Gy + Fz Gy ⎥ j +⎢ ∂y ∂z ⎣ ∂x ⎦ ⎡∂ ⎤G ∂ ∂ Fy Gz + ( Fz Gz ) ⎥ k + ⎢ ( Fx Gz ) + ∂y ∂z ⎣ ∂x ⎦
G GG ⎡∂ ∇ ⋅ FG = ⎢ ⎣ ∂x
( )
(
⎡ Fx Gx ∂ ⎤⎢ ⎥ Fy Gx ∂z ⎦ ⎢⎢ ⎣ Fz Gx
∂ ∂y
)
Fx G y
(
)
(
)
(
)
(
(1)
)
We use the product rule on each term in Eq. 1 and rearrange to get
Left hand side: G GG ⎡ ⎛ ∂F ∂Fy ∂Fz ⎞ ⎛ ∂ ∂ ∂ ⎞ ⎤G ∇ ⋅ FG = ⎢Gx ⎜ x + + + Fy + Fz ⎟ Gx ⎥ i ⎟ + ⎜ Fx ∂y ∂z ⎠ ⎝ ∂x ∂y ∂z ⎠ ⎥⎦ ⎣⎢ ⎝ ∂x ⎡ ⎛ ∂F ∂Fy ∂Fz ⎞ ⎛ ∂ ∂ ∂ ⎞ ⎤G + ⎢G y ⎜ x + + + Fy + Fz ⎟ Gy ⎥ j ⎟ + ⎜ Fx ∂y ∂z ⎠ ⎝ ∂x ∂y ∂z ⎠ ⎥⎦ ⎢⎣ ⎝ ∂x ⎡ ⎛ ∂F ∂Fy ∂Fz ⎞ ⎛ ∂ ∂ ∂ ⎞ ⎤G + ⎢ Gz ⎜ x + + + Fy + Fz ⎟ Gz ⎥ k ⎟ + ⎜ Fx ∂y ∂z ⎠ ⎝ ∂x ∂y ∂z ⎠ ⎥⎦ ⎢⎣ ⎝ ∂x
( )
We recognize that
(2)
G G ∂Fx ∂Fy ∂Fz G G ∂ ∂ ∂ + + = ∇ ⋅ F and Fx + Fy + Fz = F ⋅ ∇ . Eq. 2 then becomes ∂x ∂y ∂z ∂x ∂y ∂z
Left hand side: G GG G G G G G G G G G G ∇ ⋅ FG = ⎡Gx ∇ ⋅ F + F ⋅∇ Gx ⎤ i + ⎡G y ∇ ⋅ F + F ⋅∇ G y ⎤ j ⎣ ⎦ ⎣ ⎦ G G G G G + ⎡Gz ∇ ⋅ F + F ⋅∇ Gz ⎤ k ⎣ ⎦
( )
(
(
) ( ) ) ( )
(
) (
)
(3)
After rearrangement, Eq. 3 becomes
Left hand side: G G G G G GG G G G G G G ∇ ⋅ FG = Gx i + G y j + Gz k ∇ ⋅ F + F ⋅∇ Gx i + G y j + Gz k
( ) (
)(
) (
)(
)
(4)
G Finally, recognizing vector G twice in Eq. 4, we see that the left hand side of the given equation is identical to the right hand side, and the given equation is verified. GG GG Discussion It may seem surprising, but FG ≠ GF .
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Chapter 9 Differential Analysis of Fluid Flow 9-12 Solution
We are to prove the equation.
Analysis
G G G G We let F = ρV and G = V . Using Eq. 1 of the previous problem, we have G GG GG G G G G ∇ ⋅ ρVV = V ∇ ⋅ ρV + ρV ⋅∇ V
(
)
( ) (
)
(1)
However, since the density is not operated on in the second term of Eq. 1, it can be brought outside of the parenthesis, even though it is not a constant in general. Equation 1 can thus be written as G GG GG G G G G ∇ ⋅ ρVV = V ∇ ⋅ ρV + ρ V ⋅∇ V (2)
(
)
( ) (
)
Discussion
Equation 2 was used in this chapter in the derivation of the alternative form of Cauchy’s equation.
9-13 Solution
We are to transform cylindrical velocity components to Cartesian velocity components.
Analysis also θ,
We apply trigonometry, recognizing that the angle between u and ur is θ, and the angle between v and uθ is
x component of velocity:
u = ur cos θ − uθ sin θ
(1)
v = ur sin θ + uθ cosθ
(2)
w = uz
(3)
Similarly,
y component of velocity: The transformation of the z component is trivial,
z component of velocity: Discussion
These transformations come in handy.
9-14 Solution
We are to transform Cartesian velocity components to cylindrical velocity components.
Analysis also θ,
We apply trigonometry, recognizing that the angle between u and ur is θ, and the angle between v and uθ is
ur component of velocity:
ur = u cos θ + v sin θ
(1)
uθ = −u sin θ + v cosθ
(2)
uz = w
(3)
Similarly,
uθ component of velocity:: The transformation of the z component is trivial,
z component of velocity: Discussion
You can also obtain Eqs. 1 and 2 by solving Eqs. 1 and 2 of the previous problem simultaneously.
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Chapter 9 Differential Analysis of Fluid Flow 9-15 Solution We are to transform a given set of Cartesian coordinates and velocity components into cylindrical coordinates and velocity components.
First we apply the coordinate transformations given in this chapter,
Analysis
r = x2 + y 2 =
θ = tan −1
( 0.50 m ) + ( 0.20 m ) 2
2
= 0.5385 m
(1)
y ⎛ 0.20 m ⎞ o = tan −1 ⎜ ⎟ = 21.80 = 0.3805 radians x ⎝ 0.50 m ⎠
(2)
Next we apply the results of the previous problem, m 0.50 m m 0.20 m m × − 5.6 × = 7.484 s 0.5385 m s 0.5385 m s
(3)
m 0.20 m m 0.50 m m × − 5.6 × = −9.025 s 0.5385 m s 0.5385 m s
(4)
ur = u cos θ + v sin θ = 10.3
uθ = −u sin θ + v cosθ = −10.3
Note that we have used the fact that x = rcosθ and y = rsinθ for convenience in Eqs. 3 and 4. Our final results are summarized to three significant digits:
r = 0.539 m, θ = 0.381 radians, ur = 7.48
Results:
m m , uθ = −9.03 s s
(5)
We verify our result by calculating the square of the speed in both coordinate systems. In Cartesian coordinates, 2
2
m⎞ ⎛ m⎞ m ⎛ V 2 = u 2 + v 2 = ⎜ 10.3 ⎟ + ⎜ −5.6 ⎟ = 137.5 2 s ⎠ ⎝ s ⎠ s ⎝
2
(6)
In cylindrical coordinates, 2
2
m⎞ ⎛ m⎞ m ⎛ V 2 = ur 2 + uθ 2 = ⎜ 7.484 ⎟ + ⎜ −9.025 ⎟ = 137.5 2 s ⎠ ⎝ s ⎠ s ⎝
Discussion
2
(7)
Such checks of our algebra are always wise.
9-16 Solution We are to transform a given set of Cartesian velocity components into cylindrical velocity components, and identify the flow. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y or r-θ plane.
Analysis We recognize that r 2 = x 2 + y 2 . We also know that y = rsinθ and x = rcosθ. Using the results of Problem 914, the cylindrical velocity components are
ur component of velocity:
uθ component of velocity::
ur = u cosθ + v sin θ =
Cr sin θ cosθ Cr sin θ cosθ − =0 r2 r2
uθ = −u sin θ + v cos θ = −
Cr sin 2 θ Cr cos 2 θ −C − = r r2 r2
(1)
(2)
where we have also used the fact that cos2θ + sin2θ = 1. We recognize the velocity components of Eqs. 1 and 2 as those of a line vortex. Discussion
The negative sign in Eq. 2 indicates that this vortex is in the clockwise direction.
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Chapter 9 Differential Analysis of Fluid Flow 9-17 Solution
We are to transform a given set of cylindrical velocity components into Cartesian velocity components.
Analysis
We apply the coordinate transformations given in this chapter, along with the results of Problem 9-13, u = ur cos θ − uθ sin θ =
x component of velocity:
m x Γ y − 2π r r 2π r r
(1)
We recognize that r 2 = x 2 + y 2 . Thus, Eq. 1 becomes u=
x component of velocity:
1
(
2π x 2 + y 2
)
( mx − Γy )
(2)
Similarly,
v = ur sin θ + uθ cos θ =
y component of velocity:
m y Γ x + 2π r r 2π r r
(3)
Again recognizing that r 2 = x 2 + y 2 , Eq. 3 becomes v=
y component of velocity:
(
1
2π x 2 + y 2
)
( my + Γx )
(4)
We verify our result by calculating the square of the speed in both coordinate systems. In Cartesian coordinates, V 2 = u 2 + v2 =
1 4π
2
(
x +y 2
2
)
(m x 2
2
2
)
− 2mxΓy + Γ 2 y 2 +
1 4π
(
2
x2 + y 2
)
(m y 2
2
2
+ 2myΓx + Γ 2 x 2
)
Two of the terms in Eq. 5 cancel, and we combine the others. After simplification, Magnitude of velocity squared:
V 2 = u 2 + v2 =
1
4π
2
(x
2
+y
2
)
(m
2
+ Γ2
)
(6)
We calculate V2 from the components given in cylindrical coordinates as well, Magnitude of velocity squared:
V 2 = ur 2 + uθ 2 =
m2 Γ2 m2 + Γ2 + 2 2 = 2 2 4π r 4π r 4π 2 r 2
(7)
Finally, since r 2 = x 2 + y 2 , Eqs. 6 and 7 are the same, and the results are verified. Discussion
Such checks of our algebra are always wise.
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(5)
Chapter 9 Differential Analysis of Fluid Flow 9-18E Solution We are to transform a given set of Cartesian coordinates and velocity components into cylindrical coordinates and velocity components.
First we apply the coordinate transformations given in this chapter,
Analysis
⎛ 1 ft ⎞ x = r cos θ = 6.20 in × cos 30.0o ⎜ ⎟ = 0.4474 ft ⎝ 12 in ⎠
(1)
⎛ 1 ft ⎞ y = r sin θ = 6.20 in × sin 30.0o ⎜ ⎟ = 0.2583 ft ⎝ 12 in ⎠
(2)
(
)
and
(
)
Next we apply the results of Problem 9-13, u = ur cos θ − uθ sin θ = 1.37
(
)
(
)
ft ft ft × cos 30.0o − 3.82 × sin 30.0o = −0.7235 s s s
(3)
and v = ur sin θ + uθ cosθ = 1.37
(
)
(
)
ft ft ft × sin 30.0o + 3.82 × cos 30.0o = 3.993 s s s
(4)
Our final results are summarized to three significant digits: x = 0.447 ft , y = 0.258 ft , u = -0.724
Results:
ft ft , v = 3.99 s s
(5)
We verify our result by calculating the square of the speed in both coordinate systems. In Cartesian coordinates, 2
2
ft ⎞ ⎛ ft ⎞ ft 2 ⎛ V 2 = u 2 + v 2 = ⎜ −0.7235 ⎟ + ⎜ 3.993 ⎟ = 16.47 2 s s⎠ ⎝ s⎠ ⎝
(6)
In cylindrical coordinates, 2
2
ft ⎞ ⎛ ft ⎞ ft ⎛ V 2 = ur 2 + uθ 2 = ⎜ 1.37 ⎟ + ⎜ 3.82 ⎟ = 16.47 2 s s s ⎝ ⎠ ⎝ ⎠
Discussion
2
(7)
Such checks of our algebra are always wise.
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Chapter 9 Differential Analysis of Fluid Flow Continuity Equation
9-19C Solution
We are to discuss the material derivative of density for the case of compressible and incompressible flow.
Analysis If the flow field is compressible, we expect that as a fluid particle (a material element) moves around in the flow, its density changes. Thus the material derivative of density (the rate of change of density following a fluid particle) is non-zero for compressible flow. However, if the flow field is incompressible, the density remains constant. As a fluid particle moves around in the flow, the material derivative of density must be zero for incompressible flow (no change in density following the fluid particle). Discussion
The material derivative of any property is the rate of change of that property following a fluid particle.
9-20C Solution We are to explain why the derivation of the continuity via the divergence theorem is so much less involved than the derivation of the same equation by summation of mass flow rates through each face of an infinitesimal control volume. Analysis In the derivation using the divergence theorem, we begin with the control volume form of conservation of mass, and simply apply the divergence theorem. The control volume form was already derived in Chap. 5, so we begin the derivation in this chapter with an established conservation of mass equation. On the other hand, the alternative derivation is from “scratch” and therefore requires much more algebra. Discussion The bottom line is that the divergence theorem enables us to quickly convert the control volume form of the conservation law into the differential form.
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Chapter 9 Differential Analysis of Fluid Flow 9-21 Solution For given velocity component u and density ρ, we are to predict velocity component v, plot an approximate shape of the duct, and predict its height at section (2). Assumptions 1 The flow is steady and two-dimensional in the x-y plane, but compressible. 2 Friction on the walls is ignored. 3 Axial velocity u and density ρ vary linearly with x. 4 The x axis is a line of top-bottom symmetry. Properties
The fluid is standard air. The speed of sound is about 340 m/s, so the flow is subsonic, but compressible.
Analysis
(a) We write expressions for u and ρ, forcing them to be linear in x, u −u Cu = 2 1 = Δx
u = u1 + Cu x
ρ = ρ1 + Cρ x
Cρ =
ρ 2 − ρ1 Δx
=
m s = −100 1 2.0 m s
(1)
kg m3 = 0.175 kg 2.0 m m4
(2)
(100 − 300 )
(1.2 − 0.85 )
where Cu and Cρ are constants. We use the compressible form of the steady continuity equation, placing the unknown term v on the left hand side, and plugging in Eqs. 1 and 2, ∂ ( ρv) ∂y
=−
∂ ( ρu ) ∂x
=−
((
)
∂ ρ1 + C ρ x ( u1 + Cu x )
)
∂x
After some algebra, ∂ ( ρv)
(
)
= − ρ1Cu + u1C ρ − 2Cu C ρ x
(3)
ρ v = − ( ρ1Cu + u1Cρ ) y − 2Cu Cρ xy + f ( x )
(4)
∂y
We integrate Eq. 3 with respect to y,
Since this is a partial integration, we add an arbitrary function of x instead of simply a constant of integration. We now apply boundary conditions. Since the flow is symmetric about the x axis (y = 0), v must equal zero at y = 0 for any x. This is possible only if f(x) is identically zero. Applying f(x) = 0, dividing by ρ to solve for v, and plugging in Eq. 2, Eq. 4 becomes v=
(
)
− ρ1Cu + u1C ρ y − 2Cu Cρ xy
ρ
=
(
)
− ρ1Cu + u1Cρ y − 2Cu C ρ xy
ρ1 + C ρ x
2
1.5
(5)
Top wall
y 1
(b) For known values of u and v, we can plot streamlines between x = 0 and x = 2.0 m using the technique described in Chap. 4. Several streamlines are shown in Fig. 1. The streamline starting at x = 0, y = 0.8 m is the top wall of the duct. (c) At section (2), the top streamline crosses y = 1.70 m at x = 2.0 m. Thus, the predicted height of the duct at section (2) is 1.70 m. Discussion You can verify that the combination of Eqs. 1, 2, and 5 satisfies the steady compressible continuity equation. However, this alone does not guarantee that the density and velocity components will actually follow these equations if this diverging duct were to be built. The actual flow depends on the pressure rise between sections (1) and (2) – only one unique pressure rise can yield the desired flow deceleration. Temperature may also change considerably in this kind of compressible flow field.
0.5
0 0
0.5
1
1.5
2
x Symmetry line
FIGURE 1 Streamlines for a diverging duct.
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Chapter 9 Differential Analysis of Fluid Flow 9-22 Solution
We are to repeat Example 9-1, but without using continuity.
Assumptions 1 Density varies with time, but not space; in other words, the density is uniform throughout the cylinder at any given time, but changes with time. 2 No mass escapes from the cylinder during the compression. Analysis The mass inside the cylinder is constant, but the volume decreases linearly as the piston moves up. At t = 0 when L = LBottom the initial volume of the cylinder is V(0) = LBottomA, where A is the cross-sectional area of the cylinder. At t = 0 the density is ρ = ρ(0) = m/V(0), and thus m = ρ ( 0 ) V ( 0 ) = ρ ( 0 ) LBottom A
Mass in the cylinder:
(1)
Mass m (Eq. 1) is a constant since no mass escapes during the compression. At some later time t, L = LBottom − VP t and the volume is thus V = ( LBottom − VP t ) A
Cylinder volume at time t:
(2)
The density at time t is
ρ=
Density at time t:
ρ ( 0 ) LBottom A m = V ( LBottom − VP t ) A
(3)
where we have plugged in Eq. 1 for m and Eq. 2 for V. Equation 3 reduces to
ρ = ρ (0)
LBottom LBottom − VP t
(4)
or, using the nondimensional variables of Example 9-1, Nondimensional result:
ρ 1 = ρ (0) 1 − VP t
or
ρ* =
1 1− t *
(5)
LBottom
which is identical to Eq. 5 of Example 9-1. Discussion
We see by this exercise that the continuity equation is indeed an equation of conservation of mass.
9-23 Solution
We are to expand the continuity equation in Cartesian coordinates.
G ⎛ ∂ G ∂ G ∂ G⎞ We expand the second term by taking the dot product of the del operator ∇ = ⎜ i + j + k ⎟ with ∂y ∂z ⎠ ⎝ ∂x G G G G ρV = ( ρ u ) i + ( ρ v ) j + ( ρ w ) k , giving
Analysis
Compressible continuity equation in Cartesian coordinates: ∂ρ ∂ ( ρ u ) ∂ ( ρ v ) ∂ ( ρ w ) + + + =0 ∂t ∂x ∂y ∂z
(1)
We can further expand Eq. 1 by using the product rule on the spatial derivatives, resulting in 7 terms, Further expansion: Discussion
∂ρ ∂u ∂ρ ∂v ∂ρ ∂w ∂ρ +ρ +u +ρ +v +ρ +w =0 ∂t ∂x ∂x ∂y ∂y ∂z ∂z
(2)
We can do a similar thing in cylindrical coordinates, but the algebra is somewhat more complicated.
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Chapter 9 Differential Analysis of Fluid Flow 9-24 Solution
We are to write the given equation as a word equation and discuss it.
Analysis Here is a word equation: “The time rate of change of volume of a fluid particle per unit volume is equal to the divergence of the velocity field.” As a fluid particle moves around in a compressible flow, it can distort, rotate, and get larger or smaller. Thus the volume of the fluid element can change with time; this is represented by the left hand side of the equation. The right hand side is identically zero for an incompressible flow, but it is not zero for a compressible flow. Thus we can think of the volumetric strain rate as a measure of compressibility of a fluid flow. Discussion Volumetric strain rate is a kinematic property as discussed in Chap. 4. Nevertheless, it is shown here to be related to the continuity equation (conservation of mass).
9-25 Solution We are to verify that a given flow field satisfies the continuity equation, and we are to discuss conservation of mass at the origin. Analysis
The 2-D cylindrical velocity components (ur,uθ) for this flow field are ur =
Cylindrical velocity components:
m 2π r
uθ =
Γ 2π r
(1)
where m and Γ are constants We plug Eq. 1 into the incompressible continuity equation in cylindrical coordinates, Incompressible continuity: 1 ∂ ( rur ) 1 ∂ ( uθ ) ∂ ( u z ) + + =0 r ∂r r ∂θ ∂z
or
⎛ Γ ⎞ ⎛ m ⎞ ∂ ∂ 1 ⎜⎝ 2π ⎟⎠ 1 ⎜⎝ 2π r ⎟⎠ ∂ ( u z ) + =0 + ∂z θ
r ∂r r ∂ 0
0
(2)
0
The first term is zero because it is the derivative of a constant. The second term is zero because r is not a function of θ. The third term is zero since this is a 2-D flow with uz = 0. Thus, we verify that the incompressible continuity equation is satisfied for the given velocity field. At the origin, both ur and uθ go to infinity. Conservation of mass is not affected by uθ, but the fact that ur is nonzero at the origin violates conservation of mass. We think of the flow along the z axis as a line sink toward which mass approaches from all directions in the plane and then disappears (like a black hole in two dimensions). Mass is not conserved at the origin. Discussion Singularities such as this are unphysical of course, but are nevertheless useful as approximations of real flows, as long as we stay away from the singularity itself.
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Chapter 9 Differential Analysis of Fluid Flow 9-26 Solution
We are to verify that a given velocity field satisfies continuity.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis
The velocity field of Problem 9-16 is u=
Cartesian velocity components:
Cy x2 + y 2
v=
−Cx x2 + y2
(1)
We check continuity, staying in Cartesian coordinates, ∂u ∂x N
(
−2 xCy x 2 + y 2
∂v ∂y N
+
)
−3
(
2 yCx x 2 + y 2
+
)
−3
∂w =0 ∂z N 0 since 2-D
So we see that the incompressible continuity equation is indeed satisfied. Discussion The fact that the flow field satisfies continuity does not guarantee that a corresponding pressure field exists that can satisfy the steady conservation of momentum equation. In this case, however, it does.
9-27 Solution
We are to verify that a given velocity field is incompressible.
Assumptions
1 The flow is two-dimensional, implying no z component of velocity and no variation of u or v with z.
Analysis
The components of velocity in the x and y directions respectively are u = 1.3 + 2.8 x
v = 1.5 − 2.8 y
To check if the flow is incompressible, we see if the incompressible continuity equation is satisfied: ∂u ∂v ∂w + + =0 ∂ x ∂ y ∂z N N N 2.8
−2.8
or
2.8 − 2.8 = 0
0 since 2-D
So we see that the incompressible continuity equation is indeed satisfied. Hence the flow field is incompressible. Discussion The fact that the flow field satisfies continuity does not guarantee that a corresponding pressure field exists that can satisfy the steady conservation of momentum equation.
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Chapter 9 Differential Analysis of Fluid Flow 9-28 Solution We are to find the most general form of the radial velocity component of a purely radial flow that does not violate conservation of mass. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y or r-θ plane.
Analysis equation,
We use cylindrical coordinates for convenience. We solve for ur using the incompressible continuity ∂ ( uz ) 1 ∂ ( rur ) 1 ∂ ( uθ ) =0 + + r ∂r r ∂θ
∂z 0 for radial flow
or
∂ ( rur ) ∂r
=0
(1)
0 for 2-D flow
We integrate Eq. 1 with respect to r, adding a function of the other variable θ rather than simply a constant of integration since this is a partial integration, rur = f (θ )
Result:
or
ur =
f (θ ) r
(2)
Discussion
Any function of θ in Eq. 2 will satisfy the continuity equation.
9-29 Solution
We are to determine a relationship between constants a, b, c, and d that ensures incompressibility.
Assumptions
1 The flow is steady. 2 The flow is incompressible (under certain restraints to be determined).
Analysis
We plug the given velocity components into the incompressible continuity equation,
Condition for incompressibility:
∂u ∂v ∂w =0 + + ∂x N ∂y N ∂z N ay 2
3cy 2
ay 2 + 3cy 2 = 0
0
Thus to guarantee incompressibility, constants a and c must satisfy the following relationship: Condition for incompressibility:
a = −3c
(1)
Discussion If Eq. 1 were not satisfied, the given velocity field might still represent a valid flow field, but density would have to vary with location in the flow field – in other words the flow would be compressible.
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Chapter 9 Differential Analysis of Fluid Flow 9-30 Solution
We are to find the y component of velocity, v, using a given expression for u.
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis Since the flow is steady and incompressible, we apply the incompressible continuity in Cartesian coordinates to the flow field, giving
Condition for incompressibility:
∂v ∂u ∂w =− − ∂y ∂x N ∂z N a
∂v = −a ∂y
0
Next we integrate with respect to y. Note that since the integration is a partial integration, we must add some arbitrary function of x instead of simply a constant of integration. v = − ay + f ( x )
Solution:
If the flow were three-dimensional, we would add a function of x and z instead. Discussion To satisfy the incompressible continuity equation, any function of x will work since there are no derivatives of v with respect to x in the continuity equation. Not all functions of x are necessarily physically possible, however, since the flow must also satisfy the steady conservation of momentum equation.
9-31 Solution We are to find the most general form of the tangential velocity component of a purely circular flow that does not violate conservation of mass. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y or r-θ plane.
Analysis equation,
We use cylindrical coordinates for convenience. We solve for uθ using the incompressible continuity ∂ ( uz ) 1 ∂ ( rur ) 1 ∂ ( uθ ) =0 + + r ∂r r ∂θ ∂z
0 for circular flow
or
∂ ( uθ ) ∂θ
=0
(1)
0 for 2-D flow
We integrate Eq. 1 with respect to θ, adding a function of the other variable r rather than simply a constant of integration since this is a partial integration, Result: Discussion
uθ = f ( r )
(2)
Any function of r in Eq. 2 will satisfy the continuity equation.
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Chapter 9 Differential Analysis of Fluid Flow 9-32 Solution
We are to find the y component of velocity, v, using a given expression for u.
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis
We plug the velocity components into the steady incompressible continuity equation,
Condition for incompressibility:
∂v = −a ∂y
∂v ∂u ∂w =− − ∂y ∂x N ∂z N a
0
Next we integrate with respect to y. Note that since the integration is a partial integration, we must add some arbitrary function of x instead of simply a constant of integration.
v = −ay + f ( x )
Solution:
If the flow were three-dimensional, we would add a function of x and z instead. Discussion To satisfy the incompressible continuity equation, any function of x will work since there are no derivatives of v with respect to x in the continuity equation. Not all functions of x are necessarily physically possible, however, since the flow may not be able to satisfy the steady conservation of momentum equation.
9-33 Solution
We are to find the y component of velocity, v, using a given expression for u.
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis
We plug the velocity components into the steady incompressible continuity equation,
Condition for incompressibility:
∂v ∂u ∂w =− − x ∂y ∂ ∂z N N 2 ax − by
∂v = −2ax + by ∂y
0
Next we integrate with respect to y. Note that since the integration is a partial integration, we must add some arbitrary function of x instead of simply a constant of integration.
Solution:
v = −2axy +
by 2 + f ( x) 2
If the flow were three-dimensional, we would add a function of x and z instead. Discussion To satisfy the incompressible continuity equation, any function of x will work since there are no derivatives of v with respect to x in the continuity equation.
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Chapter 9 Differential Analysis of Fluid Flow 9-34 Solution component.
For a given axial velocity component in an axisymmetric flow field, we are to generate the radial velocity
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is axisymmetric implying that uθ = 0 and there is no variation in the θ direction. Analysis We use the incompressible continuity equation in cylindrical coordinates, simplified as follows for axisymmetric flow, 1 ∂ ( rur ) ∂ ( u z ) + =0 r ∂r ∂z
Incompressible axisymmetric continuity equation:
(1)
We rearrange Eq. 1, ∂ ( rur ) ∂r
= −r
∂ ( uz ) ∂z
= −r
u z ,exit − u z ,entrance L
(2)
We integrate Eq. 2 with respect to r, rur = −
r 2 u z ,exit − u z ,entrance + f ( z) L 2
(3)
Notice that since we performed a partial integration with respect to r, we add a function of the other variable z rather than simply a constant of integration. We divide all terms in Eq. 3 by r and recognize that the term with f(z) will go to infinity at the centerline of the nozzle (r = 0) unless f(z) = 0. We write our final expression for ur,
ur = −
Radial velocity component:
r u z ,exit − u z ,entrance 2 L
(4)
Discussion
You should plug the given equation and Eq. 4 into Eq. 1 to verify that the result is correct. (It is.)
9-35 Solution
We are to find the z component of velocity using given expressions for u and v.
Assumptions
1 The flow is steady. 2 The flow is incompressible.
Analysis
We apply the steady incompressible continuity equation to the given flow field,
Condition for incompressibility:
∂w ∂u ∂v =− − ∂z ∂x N ∂y N a + by
∂w = −a − by + bz 2 ∂z
− bz 2
Next we integrate with respect to z. Note that since the integration is a partial integration, we must add some arbitrary function of x and y instead of simply a constant of integration.
Solution:
w = − az − byz +
bz 3 + f ( x, y ) 3
Discussion To satisfy the incompressible continuity equation, any function of x and y will work since there are no derivatives of w with respect to x or y in the continuity equation.
9-19 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 9 Differential Analysis of Fluid Flow Stream Function
9-36C Solution useful.
We are to discuss the significance of curves of constant stream function, and why the stream function is
Analysis Curves of constant stream function represent streamlines of a flow. A stream function is useful because by drawing curves of constant ψ, we can visualize the instantaneous velocity field. In addition, the change in the value of ψ from one streamline to another is equal to the volume flow rate per unit width between the two streamlines. Discussion
Streamlines are an instantaneous flow description, as discussed in Chap. 4.
9-37C Solution We are to discuss the restrictions on the stream function that cause it to exactly satisfy 2-D incompressible continuity, and why they are necessary. Analysis Stream function ψ must be a smooth function of x and y (or r and θ). These restrictions are necessary so that the second derivatives of ψ with respect to both variables are equal regardless of the order of differentiation. In other ∂ 2ψ ∂ 2ψ , then the 2-D incompressible continuity equation is satisfied exactly by the definition of ψ. words, if = ∂x∂y ∂y∂x Discussion If the stream function were not smooth, there would be sudden discontinuities in the velocity field as well – a physical impossibility that would violate conservation of mass.
9-38C Solution another.
We are to discuss the significance of the difference in value of stream function from one streamline to
Analysis The difference in the value of ψ from one streamline to another is equal to the volume flow rate per unit width between the two streamlines. Discussion
This fact about the stream function can be used to calculate the volume flow rate in certain applications.
9-39 Solution
For a given stream function we are to generate expressions for the velocity components.
Assumptions
1 The flow is steady. 2 The flow is two-dimensional in the r-θ plane.
Analysis
We differentiate ψ to find the velocity components in cylindrical coordinates,
Radial velocity component:
Tangential velocity component:
uθ =
⎛ a2 ⎞ 1 ∂ψ = V cos θ ⎜ 1 − 2 ⎟ r ∂θ r ⎠ ⎝
uθ = −
⎛ a2 ⎞ ∂ψ = −V sin θ ⎜ 1 + 2 ⎟ r ⎠ ∂r ⎝
Discussion The radial velocity component is zero at the cylinder surface (r = a), but the tangential velocity component is not. In other words, this approximation does not satisfy the no-slip boundary condition along the cylinder surface. See Chap. 10 for a more detailed discussion about such approximations.
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Chapter 9 Differential Analysis of Fluid Flow 9-40 Solution We are to generate an expression for the stream function along a vertical line in a given flow field, and we are to determine ψ at the top wall. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed. Analysis We start by picking one of the two definitions of the stream function (it doesn’t matter which part we choose – the solution will be identical). ∂ψ V =u= y h ∂y
(1)
Next we integrate Eq. 1 with respect to y, noting that this is a partial integration and we must add an arbitrary function of the other variable, x, rather than a simple constant of integration.
ψ=
V 2 y + g ( x) 2h
(2)
Now we choose the other part of the definition of ψ, differentiate Eq. 2, and rearrange as follows: v=−
∂ψ = −g′( x) ∂x
(3)
where g′(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, the given equation and Eq. 3. We equate these and integrate with respect to x, we find g(x), v = 0 = −g′( x)
g′( x) = 0
g ( x) = C
(4)
Note that here we have added an arbitrary constant of integration C since g is a function of x only. Finally, plugging Eq. 4 into Eq. 2 yields the final expression for ψ,
Stream function:
ψ=
V 2 y +C 2h
(5)
We find constant C by employing the boundary condition on ψ. Here, ψ = 0 along y = 0 (the bottom wall). Thus C is equal to zero by Eq. 5, and
Stream function:
ψ=
V 2 y 2h
(6)
Along the top wall, y = h, and thus
Stream function along top wall:
ψ top =
V 2 Vh h = 2h 2
(7)
Discussion The stream function of Eq. 6 is valid not only along the vertical dashed line of the figure provided in the problem statement, but everywhere in the flow since the flow is fully developed and there is nothing special about any particular x location.
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Chapter 9 Differential Analysis of Fluid Flow 9-41 Solution We are to generate an expression for the volume flow rate per unit width for Couette flow. We are to compare results from two methods of calculation. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed. Analysis
We integrate the x component of velocity times cross-sectional area to obtain volume flow rate, y=h
y=h V ⎡ Vy 2 ⎤ Vh V = ∫ udA = ∫ yWdy = ⎢ W⎥ = W y =0 h 2 ⎣ 2h ⎦ y = 0 A
(1)
where W is the width of the channel into the page. On a per unit width basis, we divide Eq. 1 by W to get V Vh = 2 W
Volume flow rate per unit width:
(2)
The volume flow rate per unit width between any two streamlines ψ2 and ψ1 is equal to ψ2 – ψ1. We take the streamlines representing the top wall and the bottom wall of the channel. Using the result from the previous problem, V Vh Vh = ψ top −ψ bottom = −0 = W 2 2
Volume flow rate per unit width:
(3)
Equations 2 and 3 agree, as they must. Discussion
The integration of Eq. 1 can be performed at any x location in the channel since the flow is fully developed.
9-42E Solution streamlines.
We are to plot several streamlines using evenly spaced values of ψ and discuss the spacing between the
ψ = 0.50
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed.
1.2
Analysis The stream function is obtained from the result of Problem 9-40,
1
Stream function:
ψ=
V 2 y 2h
(1)
We solve Eq. 1 for y as a function of ψ so that we can plot streamlines,
Equation for streamlines:
y=
2hψ V
(2)
We have taken only the positive root in Eq. 2 for obvious reasons. Along the top wall, y = h, and thus
ψ top
ft 10.0 × 0.100 ft Vh ft 2 s = = = 0.500 2 2 s
y (in)
0.8 0.6 0.4
ψ = 0.10 ψ = 0.05
0.2
ψ=0 0
(3)
0
0.2
0.4
0.6
0.8
1
x (in)
The streamlines themselves are straight, flat horizontal lines as seen by Eq. 1. We divide ψtop by 10 to generate evenly spaced stream functions. We plot 11 streamlines in the figure (counting the streamlines on both walls) by plugging these values of ψ into Eq. 2. The streamlines are not evenly spaced. This is because the volume flow rate per unit width between two streamlines ψ2 and ψ1 is equal to ψ2 – ψ1. The flow speeds near the top of the channel are higher than those near the bottom of the channel, so we expect the streamlines to be closer near the top. Discussion The extent of the x axis in the figure is arbitrary since the flow is fully developed. You can immediately see from a streamline plot like Fig. 1 where flow speeds are high and low (relatively speaking).
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Chapter 9 Differential Analysis of Fluid Flow 9-43 Solution
We are to generate an expression for the stream function along a vertical line in a given flow field.
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed. Analysis We start by picking one of the two definitions of the stream function (it doesn’t matter which part we choose – the solution will be identical). ∂ψ 1 dP 2 y − hy =u = 2 μ dx ∂y
(
)
(1)
Next we integrate Eq. 1 with respect to y, noting that this is a partial integration and we must add an arbitrary function of the other variable, x, rather than a simple constant of integration.
ψ=
1 dP ⎛ y 3 y2 ⎞ ⎜ − h ⎟ + g ( x) 2 μ dx ⎝ 3 2 ⎠
(2)
Now we choose the other part of the definition of ψ, differentiate Eq. 2, and rearrange as follows:
v=−
∂ψ = −g′( x) ∂x
(3)
where g′(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, the given equation and Eq. 3. We equate these and integrate with respect to x to find g(x), v = 0 = −g′( x)
g′( x) = 0
g ( x) = C
(4)
Note that here we have added an arbitrary constant of integration C since g is a function of x only. Finally, plugging Eq. 4 into Eq. 2 yields the final expression for ψ,
Stream function:
ψ=
1 dP ⎛ y 3 y2 ⎞ − h ⎜ ⎟+C 2 μ dx ⎝ 3 2 ⎠
(5)
We find constant C by employing the boundary condition on ψ. Here, ψ = 0 along y = 0 (the bottom wall). Thus C is equal to zero by Eq. 5, and
Stream function:
ψ=
1 dP ⎛ y 3 y2 ⎞ ⎜ −h ⎟ 2 μ dx ⎝ 3 2 ⎠
(6)
Along the top wall, y = h, and thus
Stream function along top wall:
ψ top = −
1 dP 3 h 12 μ dx
(7)
Discussion The stream function of Eq. 6 is valid not only along the vertical dashed line of the figure provided in the problem statement, but everywhere in the flow since the flow is fully developed and there is nothing special about any particular x location.
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Chapter 9 Differential Analysis of Fluid Flow 9-44 Solution We are to generate an expression for the volume flow rate per unit width for fully developed channel flow. We are to compare results from two methods of calculation. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed. Analysis
We integrate the x component of velocity times cross-sectional area to obtain volume flow rate, y=h
⎡ 1 dP ⎛ y 3 y = h 1 dP y2 ⎞ ⎤ − y 2 − hy Wdy = ⎢ h V = ∫ udA = ∫ ⎜ ⎟W ⎥ y = 0 2 μ dx 2 ⎠ ⎦⎥ y = 0 A ⎣⎢ 2 μ dx ⎝ 3
(
)
1 dP ⎛ h3 ⎞ 1 dP 3 hW = ⎜ − ⎟W = − 2 μ dx ⎝ 6 ⎠ 12 μ dx
(1)
where W is the width of the channel into the page. On a per unit width basis, we divide Eq. 1 by W to get
Volume flow rate per unit width:
V 1 dP 3 =− h W 12μ dx
(2)
The volume flow rate per unit width between any two streamlines ψ2 and ψ1 is equal to ψ2 – ψ1. We take the streamlines representing the top wall and the bottom wall of the channel. Using the result from the previous problem,
Volume flow rate per unit width: V 1 dP 3 1 dP 3 = ψ top −ψ bottom = − h −0 = − h W 12μ dx 12μ dx
(3)
Equations 2 and 3 agree, as they must. Discussion
The integration of Eq. 1 can be performed at any x location in the channel since the flow is fully developed.
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Chapter 9 Differential Analysis of Fluid Flow 9-45 Solution streamlines.
We are to plot several streamlines using evenly spaced values of ψ and discuss the spacing between the
ψ = ψtop = 0.002874 Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane. 4 The flow is fully developed. Properties kg/(m⋅s). Analysis 9-43,
1.2 1
The viscosity of water at T = 20oC is 1.002 × 10-3
0.8
y (mm)
The stream function is obtained from the result of Problem
Stream function:
0.6
ψ=
1 dP ⎛ y 3 y2 ⎞ ⎜ −h ⎟ 2 μ dx ⎝ 3 2 ⎠
0.4
(1)
We need to solve Eq. 1 (a cubic equation) for y as a function of ψ so that we can plot streamlines. First we re-write Eq. 1 in standard cubic form, y3 −
Standard cubic form:
3h 2 6μψ y − =0 2 dP / dx
0.2
0
(2)
(
)
0.2
0.4
0.6
0.8
1
x (mm)
We can either look up the solution for cubic equations or use Newton’s iteration method to obtain y for a given value of ψ. In general there are three roots – we choose the positive real root with 0 < y < h, which is the only one that has physical meaning for this problem. Along the top wall, y = h, and Eq. 1 yields
Stream function along top wall: 1 dP 3 1 ψ top = − h = 12 μ dx 12 1.002 × 10 −3 kg/m ⋅ s
ψ = 0.0002874
ψ=0 0
( 20, 000 N/m ) ( 0.00120 m ) 3
FIGURE 1 Streamlines for 2-D channel flow with evenly spaced values of stream function. Values of ψ are in units of m2/s.
3
⎛ kg ⋅ m ⎞ ⎜ s2 N ⎟ ⎝ ⎠
= 2.874 × 10−3 m 2 /s
The streamlines themselves are straight, flat horizontal lines as seen by Eq. 1. We divide ψtop by 10 to generate evenly spaced stream functions. We plot 11 streamlines in Fig. 1 (counting the streamlines on both walls) by plugging these values of ψ into Eq. 2 and solving for y. The streamlines are not evenly spaced. This is because the volume flow rate per unit width between two streamlines ψ2 and ψ1 is equal to ψ2 – ψ1. The flow speeds in the middle of the channel are higher than those near the top or bottom of the channel, so we expect the streamlines to be closer near the middle. Discussion The extent of the x axis in Fig. 1 is arbitrary since the flow is fully developed. You can immediately see from a streamline plot like Fig. 1 where flow speeds are high and low (relatively speaking).
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Chapter 9 Differential Analysis of Fluid Flow 9-46 Solution
We are to calculate the volume flow rate and average speed of air being sucked through a sampling probe.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional.
Analysis For 2-D incompressible flow the difference in the value of the stream function between two streamlines is equal to the volume flow rate per unit width between the two streamlines. Thus, 2
m Volume flow rate through the sampling probe: V = (ψ u −ψ l ) × W = ( 0.150 − 0.105 ) × 0.052 m = 0.00234 m 3 /s s
(1)
The average speed of air in the probe is obtained by dividing volume flow rate by cross-sectional area,
Average speed through the sampling probe:
Vavg =
V 0.00234 m3 /s = = 10.0 m/s hW ( 0.0045 m )( 0.052 m )
(2)
Discussion Notice that the streamlines inside the probe are more closely packed than are those outside the probe because the flow speed is higher inside the probe.
9-47 Solution We are to sketch streamlines for the case of a sampling probe with too little suction, and we are to name this type of sampling and label the lower and upper dividing streamlines. Analysis If the suction were too weak, the volume flow rate through the probe would be too low and the average air speed through the probe would be lower than that of the air stream. The dividing streamlines would diverge outward rather than inward as sketched in Fig. 1. We would call this type of sampling subisokinetic sampling. Discussion We have drawn the streamlines inside the probe further apart than those in the air stream because the flow speed is lower inside the probe.
ψ = ψu
Dividing streamlines Sampling probe
h
Vfreestream
Vavg
FIGURE 1 Streamlines for subisokinetic sampling.
9-48 Solution
We are to calculate the speed of the air stream of Fig. P9-46.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional.
Analysis
In the air stream far upstream of the probe,
Volume flow rate per unit width:
V
ψ = ψl
V = ψ u −ψ l = Vfreestream yu − V∞ yl = Vfreestream ( yu − yl ) W
(1)
By definition of streamlines, the volume flow rate between the two dividing streamlines must be the same as that through the probe itself. We know the volume flow rate through the probe from the results of Problem 9-46. The value of the stream function on the lower and upper dividing streamlines are the same as those of Problem 9-46, namely ψl = 0.105 m2/s and ψu = 0.150 m2/s respectively. We also know yu – yl from the information given here. Thus, Eq. 1 yields
Freestream speed:
Vfreestream =
ψ u −ψ l
( yu − yl )
=
m2 s = 7.76 m 0.0058 m s
( 0.150 − 0.105)
(2)
Discussion We verify by these calculations that the sampling is superisokinetic (average speed through the probe is higher than that of the upstream air stream).
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Chapter 9 Differential Analysis of Fluid Flow 9-49 Solution For a given velocity field we are to generate an expression for ψ, and we are to calculate the volume flow rate per unit width between two streamlines. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis We start by picking one of the two definitions of the stream function (it doesn’t matter which part we choose – the solution will be identical). ∂ψ = u =V ∂y
(1)
Next we integrate Eq. 1 with respect to y, noting that this is a partial integration and we must add an arbitrary function of the other variable, x, rather than a simple constant of integration.
ψ = Vy + g ( x )
(2)
Now we choose the other part of the definition of ψ, differentiate Eq. 2, and rearrange as follows: v=−
∂ψ = −g′( x) ∂x
(3)
where g′(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, the given equation and Eq. 3. We equate these and integrate with respect to x to find g(x), v = 0 = −g′( x)
g′( x) = 0
g ( x) = C
(4)
Note that here we have added an arbitrary constant of integration C since g is a function of x only. Finally, plugging Eq. 4 into Eq. 2 yields the final expression for ψ,
ψ = Vy + C
Stream function:
(5)
Constant C is arbitrary; it is common to set it to zero, although it can be set to any desired value. Here, ψ = 0 along the streamline at y = 0, forcing C to equal zero by Eq. 5. For the streamline at y = 0.5 m, Value of ψ2:
⎛
ψ 2 = ⎜ 8.9 ⎝
m⎞ m2 × ( 0.5 m ) = 4.45 ⎟ s ⎠ s
(6)
The volume flow rate per unit width between streamlines ψ2 and ψ0 is equal to ψ2 – ψ0,
Volume flow rate per unit width:
V m2 m2 = ψ 2 −ψ 0 = ( 4.45 − 0 ) = 4.45 s W s
(7)
We verify our result by calculating the volume flow rate per unit width from first principles. Namely, volume flow rate is equal to speed times cross-sectional area,
Volume flow rate per unit width: V m m2 = V ( y2 − y0 ) = 8.9 × ( 0.5 − 0 ) m = 4.45 W s s
(8)
Discussion If constant C were some value besides zero, we would still get the same result for the volume flow rate since C would cancel out in the subtraction.
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Chapter 9 Differential Analysis of Fluid Flow 9-50E Solution For a given velocity field we are to generate an expression for ψ and plot several streamlines for given values of constants a and b. Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis
We plug the given equation into the steady incompressible continuity equation, ∂v ∂u ∂w =− − ∂y ∂ ∂z x N N
Condition for incompressibility:
2 ax − by
∂v = −2ax + by ∂y
0
Next we integrate with respect to y. Note that since the integration is a partial integration, we must add some arbitrary function of x instead of simply a constant of integration.
v = −2axy +
y component of velocity:
by 2 + f ( x) 2
If the flow were three-dimensional, we would add a function of x and z instead. We are told that v = 0 for all values of x when y = 0. This is only possible if f(x) = 0. Thus,
v = −2axy +
y component of velocity:
by 2 2
(1)
To obtain the stream function, we start by picking one of the two parts of the definition of the stream function,
4
∂ψ = u = ax 2 − bxy ∂y
-3 -2.4 3
Next we integrate the above equation with respect to y, noting that this is a partial integration and we must add an arbitrary function of the other variable, x, rather than a simple constant of integration.
ψ = ax 2 y −
bxy 2 + g ( x) 2
-1.2 y (ft) 2
(2)
Now we choose the other part of the definition of ψ, differentiate Eq. 2, and rearrange as follows: v=−
by 2 ∂ψ = −2axy + − g′( x) ∂x 2
g ( x) = C
2.4 1.8
ψ = 0 ft2/s
(4)
ψ = ax 2 y −
bxy 2 2
3
1.2
1
(3)
Note that here we have added an arbitrary constant of integration C since g is a function of x only. But C must be zero in order for ψ to be zero for any value of x when y = 0. Finally, Eq. 2 yields the final expression for ψ,
Solution:
-0.6 0.6
where g′(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, Eq. 1 and Eq. 3. We equate these and integrate with respect to x to find g(x), g′( x) = 0
-1.8
0 0
1
2
3
x (ft)
FIGURE 1 Streamlines for a given velocity field; the value of constant ψ is indicated for each streamline in units of ft2/s.
(5)
To plot the streamlines, we note that Eq. 5 represents a family of curves, one unique curve for each value of the stream function ψ. We solve Eq. 5 for y as a function of x. A bit of algebra (the quadratic rule) yields
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Chapter 9 Differential Analysis of Fluid Flow y=
Equation for streamlines:
ax 2 ± a 2 x 4 − 2ψ bx
(6)
bx
For the given values of constants a and b, we plot Eq. 6 for several values of ψ in Fig. 1; these curves of constant ψ are streamlines of the flow. Note that both the positive and negative roots of Eq. 6 are required to plot each streamline. The direction of the flow is found by calculating u and v at some point in the flow field. We pick x = 2 ft, y = 2 ft, where u = −1.2 ft/s and v = −2.1 ft/s. This indicates flow to the lower left near this location. We fill in the rest of the arrows in Fig. 1 to be consistent. We see that the flow enters from the upper right, and splits into two parts – one to the lower right and one to the upper left. Discussion It is always a good idea to check your algebra. In this example, you should differentiate Eq. 5 to verify that the velocity components of the given equation are obtained.
9-51 Solution
For a given velocity field we are to generate an expression for ψ.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis We start by picking one of the two definitions of the stream function (it doesn’t matter which part we choose – the solution will be identical). ∂ψ = u = V cos α ∂y
(1)
Next we integrate Eq. 1 with respect to y, noting that this is a partial integration and we must add an arbitrary function of the other variable, x, rather than a simple constant of integration.
ψ = yV cos α + g ( x )
(2)
Now we choose the other part of the definition of ψ, differentiate Eq. 2, and rearrange as follows: v=−
∂ψ = −g′( x) ∂x
(3)
where g′(x) denotes dg/dx since g is a function of only one variable, x. We now have two expressions for velocity component v, the given equation and Eq. 3. We equate these and integrate with respect to x to find g(x), v = V sin α = − g ′ ( x )
g ′ ( x ) = −V sin α
g ( x ) = − xV sin α + C
(4)
Note that here we have added an arbitrary constant of integration C since g is a function of x only. Finally, plugging Eq. 4 into Eq. 2 yields the final expression for ψ,
Stream function:
ψ = V ( y cos α − x sin α ) + C
(5)
Constant C is arbitrary; it is common to set it to zero, although it can be set to any desired value. Discussion
You can verify by differentiating ψ that Eq. 5 yields the correct values of u and v.
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Chapter 9 Differential Analysis of Fluid Flow 9-52 Solution
For a given stream function, we are to calculate the velocity components and verify incompressibility.
Assumptions 1 The flow is steady. 2 The flow is incompressible (this assumption is to be verified). 3 The flow is twodimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis
(a) We use the definition of ψ to obtain expressions for u and v.
Velocity components:
u=
∂ψ = bx + 2cy ∂y
v=−
∂ψ = −2ax − by ∂x
(1)
(b) We check if the incompressible continuity equation in the x-y plane is satisfied by the velocity components of Eq. 1, ∂u ∂v ∂w + + =0 ∂x N ∂y N ∂z N
Incompressible continuity:
b
−b
b−b = 0
(2)
0
We conclude that the flow is indeed incompressible. Discussion Since ψ is a smooth function of x and y, it automatically satisfies the continuity equation by its definition. Equation 2 confirms this. If it did not, we would go back and look for an algebra mistake somewhere.
9-53 [Also solved using EES on enclosed DVD] Solution We are to plot several streamlines for a given velocity field. Analysis We re-write the stream function equation of the previous problem with all the terms on one side,
cy 2 + bxy + ax 2 − ψ = 0
(1)
For any constant value of ψ (along a streamline), Eq. 1 is in a form that enables us to use the quadratic rule to solve for y as a function of x, Equation for a streamline:
y=
(
−bx ± b 2 x 2 − 4c ax 2 − ψ
)
(2)
2c
We plot the streamlines in Fig. 1. For each value of ψ there are two curves – one for the positive root and one for the negative root of Eq. 2. There is symmetry about a diagonal line through the origin. The streamlines appear to be hyperbolae. We determine the flow direction by plugging in a couple values of x and y and calculating the velocity components; e.g., at x = 1 m and y = 3 m, u = 2.7 m/s and v = 4.9 m/s. The flow at this point is in the upper right direction. Similarly, at x = 1 m and y = -2 m, u = -3.3 m/s and v = -1.6 m/s. The flow at this point is in the lower left direction. Discussion This flow may not represent any particular physical flow field, but it produces an interesting flow pattern.
4
ψ=6 3
2
ψ=0 1
y 0 (m) -1
ψ=0 -2
-3
ψ=6 -4 -2
-1
0
1
2
x (m) FIGURE 1 Streamlines for a given velocity field. Values of ψ are in units of m2/s.
9-30 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 9 Differential Analysis of Fluid Flow 9-54 Solution
For a given stream function, we are to calculate the velocity components and verify incompressibility.
Assumptions 1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane, implying that w = 0 and neither u nor v depend on z. Analysis
(a) We use the definition of ψ to obtain expressions for u and v.
Velocity components:
u=
∂ψ = −2by + dx ∂y
v=−
∂ψ = −2ax − c − dy ∂x
(1)
(b) We check if the incompressible continuity equation in the x-y plane is satisfied by the velocity components of Eq. 1, ∂u ∂v ∂w + + =0 ∂x N ∂y N ∂z N
Incompressible continuity:
d
−d
d −d = 0
(2)
0
We conclude that the flow is indeed incompressible. Discussion Since ψ is a smooth function of x and y, it automatically satisfies the continuity equation by its definition. Eq. 2 confirms this. If it did not, we would go back and look for an algebra mistake somewhere.
9-55 Solution
We are to make up a stream function ψ(x,y), calculate the velocity components and verify incompressibility.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis Every student should have a different stream function. He or she then takes the derivatives with respect to y and x to find u and v. The student should then plug his/her velocity components into the incompressible continuity equation. Continuity will be satisfied regardless of ψ(x,y), provided that ψ(x,y) is a smooth function of x and y. Discussion definition.
As long as ψ is a smooth function of x and y, it automatically satisfies the continuity equation by its
9-56 Solution
We are to calculate the percentage of flow going through one branch of a branching duct.
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis For 2-D incompressible flow the difference in the value of the stream function between two streamlines is equal to the volume flow rate per unit width between the two streamlines. Thus,
Main branch:
V W
⎞ m2 m2 = ψ upper wall −ψ lower wall = ( 4.15 − 2.03) = 2.12 ⎟⎟ s s ⎠ main
(1)
V W
⎞ m2 m2 = ψ upper wall −ψ branch wall = ( 4.15 − 2.80 ) = 1.35 ⎟⎟ s s ⎠ upper
(2)
Similary, in the upper branch, Upper branch:
On a percentage basis, the percentage of volume flow through the upper branch is calculated as V ⎞ m2 ⎟ Vupper W ⎟⎠ upper 1.35 s = = = 0.637 = 63.7% m2 Vmain V ⎞ 2.12 ⎟ s W ⎟⎠ main
Discussion
(3)
No dimensions are given, so it is not possible to calculate velocities.
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Chapter 9 Differential Analysis of Fluid Flow 9-57 Solution We are to calculate duct height h for a given average velocity through a duct and values of stream function along the duct walls. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis The volume flow rate through the main branch of the duct is equal to the average velocity times the crosssectional area of the duct, V = VavgWh
Volume flow rate:
(1)
We solve for h in Eq. 1, using the results of Problem 9-56, h=
Duct height:
1 V ⎞ 1 m 2 ⎛ 100 cm ⎞ = × 2.12 = 18.6 cm ⎟ m Vavg W ⎟⎠ s ⎜⎝ m ⎟⎠ 11.4 main s
(2)
An alternative way to solve for height h is to assume uniform flow in the main branch, for which ψ = Vavgy. We take the difference between ψ at the top of the duct and ψ at the bottom of the duct to find h,
ψ upper wall − ψ lower wall = Vavg yupper wall − Vavg ylower wall = Vavg ( yupper wall − ylower wall ) = Vavg h Thus, Duct height: h =
ψ upper wall − ψ lower wall Vavg
=
m2 s ⎛ 100 cm ⎞ = 18.6 cm ⎜ m ⎟ m ⎝ ⎠ 11.4 s
( 4.15 − 2.03)
(3)
You can see that we get the same result as that of Eq. 2. Discussion The result is correct even if the velocity profile through the duct is not uniform, since we have used the average velocity in our calculations.
9-58 Solution
We are to verify that the given ψ satisfies the continuity equation, and we are to discuss any restrictions.
Assumptions only).
1 The flow is steady. 2 The flow is incompressible. 3 The flow is axisymmetric (ψ is a function of r and z
Analysis
We plug the given velocity components into the axisymmetric continuity equation, ⎛ ∂ψ ⎞ ⎛ 1 ∂ψ ⎞ ∂ − ∂⎜ 1 ∂ ( rur ) ∂ ( u z ) 1 ⎜⎝ ∂z ⎟⎠ r ∂r ⎟⎠ 1 ⎛ ∂ 2ψ ∂ 2ψ ⎞ + = + ⎝ = ⎜− + ⎟=0 r ∂r r r ⎝ ∂r ∂z ∂z∂r ⎠ ∂z ∂r ∂z
Thus we see that continuity is satisfied by the given stream function. The only restriction on ψ is that ψ must be a smooth function of r and z.
Discussion
For a smooth function of two variables, the order of differentiation does not matter.
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Chapter 9 Differential Analysis of Fluid Flow 9-59 Solution We are to determine the value of the stream function along the positive y axis and the negative x axis for the case of a line source at the origin. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y or r-θ plane.
Analysis For 2-D incompressible flow the difference in the value of the stream function between two streamlines is equal to the volume flow rate per unit width between the two streamlines. Let us take the arc of the circle of radius r between the positive x axis and the positive y axis of the figure provided with the problem statement. The volume flow rate per unit width through this arc is one-fourth of V / L , the total volume flow rate per unit width, since the arc spans exactly one-fourth of the circumference of the circle.
ψ positive y axis −ψ positive x axis =
1 V 4L
(1)
Since ψ = 0 along the positive x axis, we conclude that
ψ positive y axis =
ψ along positive y axis:
1 V 4L
(2)
Similarly, the volume flow rate through the top half of the circle is half of the total volume flow rate and we conclude that
ψ negative y axis =
ψ along negative x axis:
1 V 2L
(3)
Discussion Some CFD codes use ψ as a variable, and we thus need to specify the value of ψ along boundaries of the computational domain. Simple calculations such as this are useful in these situations.
9-60 Solution We are to determine the value of the stream function along the positive y axis and the negative x axis for the case of a line sink at the origin. Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y or r-θ plane.
Analysis Everything is the same as in Problem 9-59 except that the flow direction is reversed everywhere. The volume flow rate per unit width through the arc of radius r between the positive x axis and the positive y axis of Fig. P9-59 is now negative one-fourth of V / L since the flow is now mathematically negative.
ψ positive y axis −ψ positive x axis = −
1 V 4L
(1)
Since ψ = 0 along the positive x axis, we conclude that
ψ along positive y axis:
ψ positive y axis = −
1 V 4L
(2)
Similarly, the volume flow rate through the top half of the circle is half of the total volume flow rate and is negative. We conclude that
ψ along negative x axis: Discussion
ψ negative x axis = −
1 V 2L
(3)
We need to be careful of the sign of ψ.
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Chapter 9 Differential Analysis of Fluid Flow 9-61 Solution
We are to generate an expression for the stream function that describes a given velocity field.
Assumptions only).
1 The flow is steady. 2 The flow is incompressible. 3 The flow is axisymmetric (ψ is a function of r and z
Analysis
The r and z velocity components from Problem 9-34 are
Velocity field:
ur = −
r u z ,exit − u z ,entrance 2 L
u z = u z ,entrance +
u z ,exit − u z ,entrance L
z
(1)
To generate the stream function we use the definition of ψ for steady, incompressible, axisymmetric flow, ur = −
Axisymmetric stream function:
1 ∂ψ r ∂z
uz =
1 ∂ψ r ∂r
(2)
We choose one of the definitions of Eq. 2 to integrate. We pick the second one, ⎛
ψ = ∫ ru z dr = ∫ r ⎜ u z ,entrance + Integration:
⎝
u z ,exit − u z ,entrance ⎞ z ⎟ dr L ⎠
u − u z ,entrance ⎞ r ⎛ = ⎜ u z ,entrance + z ,exit z ⎟ + f ( z) L 2⎝ ⎠ 2
(3)
We added a function of z instead of a constant of integration since this is a partial integration. Now we take the z derivative of Eq. 3 and use the other half of Eq. 2, ur = −
Differentiation:
r u z ,exit − u z ,entrance 1 1 ∂ψ =− − f ′( z ) r ∂z L r 2
(4)
We equate Eq. 4 to the known value of ur from Eq. 1, Comparison: r u z ,exit − u z ,entrance 1 r u z ,exit − u z ,entrance ur = − − f ′( z ) = − L r L 2 2
or
f ′( z ) = 0
(5)
Since f is a function of z only, integration of Eq. 5 yields f(z) = constant. The final result is thus Stream function:
ψ=
u z ,exit − u z ,entrance ⎞ r2 ⎛ z ⎟ + constant ⎜ u z ,entrance + L 2⎝ ⎠
(6)
Discussion The constant of integration can be any value since velocity components are determined by taking derivatives of the stream function.
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Chapter 9 Differential Analysis of Fluid Flow 9-62E [Also solved using EES on enclosed DVD] Solution We are to calculate the axial speed at the entrance and exit of the nozzle, and we are to plot several streamlines for a given axisymmetric flow field. Assumptions only).
1 The flow is steady. 2 The flow is incompressible. 3 The flow is axisymmetric (ψ is a function of r and z
Analysis (a) Since uz is not a function of radius, the axial velocity profile across a cross section of the nozzle is uniform. (This is consistent with the assumption that frictional effects along the nozzle walls are neglected.) Thus, at any cross section the axial speed is equal to the volume flow rate divided by cross-sectional area,
Entrance axial speed: u z ,entrance
gal 2 4 × 2.0 ⎛ 0.1337 ft 3 ⎞ ⎛ 12 in ⎞ ⎛ min ⎞ ft 4V min = = ⎜ ⎟⎜ ⎟ ⎜ 60 s ⎟ = 3.268 s gal ft π Dentrance 2 π ( 0.50 in )2 ⎝ ⎠ ⎝ ⎠ ⎠⎝
(1)
Similarly, Exit axial speed:
u z ,exit =
4V ft = 41.69 2 s π Dexit
0.25
(2)
Nozzle wall
0.2
(b) We use the stream function developed in Problem 9-61. Setting the constant to zero for simplicity, we have Stream function:
r2 ψ= 2
u z ,exit − u z ,entrance ⎛ ⎜ u z ,entrance + L ⎝
⎞ z⎟ ⎠
0.1
(3)
We solve Eq. 3 for r as a function of z and plot several streamlines in Fig. 1, Streamlines:
2ψ
r=± u z ,entrance +
0.15
0.05
r (in)
ψ=0
0
-0.05 -0.1
u z ,exit − u z ,entrance L
z
(4)
-0.15 -0.2
At the nozzle entrance (z = 0), the wall is at r = Dentrance/2 = 0.25 inches. Eq. 3 yields ψwall = 0.0007073 ft3/s for the streamline that passes through this point. This streamline thus represents the shape of the nozzle wall, and we have designed the nozzle shape.
Discussion You can verify that the diameter between the outermost streamlines varies from Dentrance to Dexit.
-0.25 0
0.5
1
1.5
2
z (in) FIGURE 1 Streamlines for flow through an axisymmetric garden hose nozzle. Note that the vertical axis is highly magnified.
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Chapter 9 Differential Analysis of Fluid Flow 9-63 Solution minimum.
We are to discuss the sign of the stream function in a separation bubble, and determine where ψ is a
Assumptions
1 The flow is steady. 2 The flow is incompressible. 3 The flow is two-dimensional in the x-y plane.
Analysis For 2-D incompressible flow the difference in the value of the stream function between two streamlines is equal to the volume flow rate per unit width between the two streamlines. For example, ψupper – ψ0 is positive and represents the volume flow rate per unit width between the wall and the uppermost streamline, The flow between these two streamlines is to the right. Likewise, the difference between ψ along the dividing streamline and ψ = ψ1 along a streamline in the upper part of the separation bubble must also be positive as sketched in Fig. 1. The arcshaped dividing streamline divides fluid within the separation bubble from fluid outside of the separation bubble. The stream function along this dividing streamline must be zero since it intersects the wall where ψ = 0. The only way we can have flow to the right in the upper part of the separation bubble is if ψ1 is negative (Fig. 1). We conclude that for this problem, all streamlines within the separation bubble have negative values of stream function. The minimum value of ψ occurs in the center of the separation bubble as sketched in Fig. 1.
ψ=0
ψ 1, viscous forces are negligible compared to inertial forces, and the viscous term in the Navier-Stokes equation can be ignored. Note that this applies only to regions outside of boundary layers, because the characteristic length scale for a boundary layer is generally much smaller than that for the overall flow. Discussion You must keep in mind that the approximations discussed here are appropriate only in certain regions of the flow field. In other regions of the same flow field, different approximations may apply.
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Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-7C Solution
We are to discuss the criterion for using modified pressure.
Analysis
Modified pressure can be used only when there are no free surface effects in the problem.
Discussion Modified pressure is simply a combination of thermodynamic pressure and hydrostatic pressure. It turns out that if there are no free surface effects, the hydrostatic pressure component is independent of the flow pressure component, and these two can be separated.
10-8C Solution
We are to discuss which nondimensional parameter is eliminated by use of the modified pressure.
Analysis Modified pressure effectively combines the effects of actual pressure and gravity. In the nondimensionalized Navier-Stokes equation in terms of modified pressure, the Froude number disappears. The reason Froude number is eliminated is because the gravity term is eliminated from the equation. Discussion
Keep in mind that we can employ modified pressure only for flows without free surface effects.
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Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-9 Solution We are to plug the given scales for this flow problem into the nondimensionalized Navier-Stokes equation to show that only two terms remain in the region consisting of most of the tank. Assumptions
1 The flow is incompressible. 2 d > d. The last term (the viscous term) is a little trickier. We know that if the flow remains laminar, the order of magnitude of Rejet is at most 103. Thus, in order for the viscous term to be of the same order of magnitude as the inertial term, d2/D2 must be of order of magnitude 10-3. Thus, provided that these criteria are met, the only two remaining terms in the Navier-Stokes equation are the pressure and gravity terms. The final dimensional form of the equation is the same as that of fluid statics, G G (6) Incompressible Navier-Stokes equation for fluid statics: ∇P = ρ g The criteria for Carrie’s approximation to be appropriate depends on the desired precision. For 1% error, D must be at least 10 times greater than d to ignore the unsteady term and the inertial term. The viscous term, however, depends on the value of Rejet. To be safe, Carrie should assume the highest possible value of Rejet, for which we know from the above order of magnitude estimates that D must be at least 103/2 times greater than d. Discussion
We cannot use the modified pressure in this problem since there is a free surface.
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Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-10 Solution We are to sketch the profile of modified pressure and G shade in the region representing hydrostatic pressure. P′ g Assumptions 1 The flow is incompressible. 2 The flow is fully developed. 3 Gravity acts vertically downward. 4 There are no free surface effects in this flow field.
P z x x1 (a) Hydrostatic pressure component
Analysis By definition, modified pressure P′ = P + ρgz. So we add hydrostatic pressure component ρgz to the given profile for P to obtain the profile for P′. Recall from Example 9-16, that for the case in which gravity does not act in the x-z plane, the pressure would be constant along any slice x = x1. Thus we infer that here (with gravity), the linear increase in P as we move down vertically in the channel is due to hydrostatic pressure. Therefore, when we add ρgz to P to obtain the modified pressure, it turns out that P′ is constant at this horizontal location.
z
x G g P P′
We show two solutions in Fig. 1: (a) datum plane z = 0 located at the bottom wall, and (b) datum plane z = 0 located at the top wall. The shaded region in Fig. 1b represents the hydrostatic pressure component. P′ is constant along the slice x = x1 for either case, and the datum plane can be drawn at any arbitrary elevation. Discussion It should be apparent why it is advantageous to use modified pressure; namely, the gravity term is eliminated from the Navier-Stokes equation, and P′ is in general simpler than P.
10-11 Solution
x1 (b)
FIGURE 1 Actual pressure P (black arrows) and modified pressure P′ (gray arrows) for fully developed planar Poiseuille flow. (a) Datum plane at bottom wall and (b) datum plane at top wall. The hydrostatic pressure component ρgz is the shaded area in (b).
We are to discuss how modified pressure varies with downstream distance in planar Poiseuille flow.
Assumptions 1 The flow is incompressible. 2 The flow is fully developed. 3 Gravity acts vertically downward. 4 There are no free surface effects in this flow field.
G g
P′ P
z
Analysis For fully developed planar Poiseuille flow between two x parallel plates, we know that pressure P decreases linearly with x, the x1 x2 distance down the channel. Modified pressure is defined as P′ = P + ρgz. However, since the flow is horizontal, elevation z does not change as we FIGURE 1 move axially down the channel. Thus we conclude that modified Actual pressure P (black arrows) and pressure P′ decreases linearly with x. We sketch both P and P′ in Fig. 1 modified pressure P′ (gray arrows) at two at two axial locations, x = x1 and x = x2. The shaded region in Fig. 1 axial locations for fully developed planar represents the hydrostatic pressure component ρgz. Since channel height Poiseuille flow. is constant, the hydrostatic component does not change with x. P′ is constant along any vertical slice, but its magnitude decreases linearly with x as sketched. Discussion The pressure gradient dP′/dx in terms of modified pressure is the same as the pressure gradient ∂P/∂x in terms of actual pressure.
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Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-12 Solution We are to generate an “exact” solution of the Navier-Stokes equation for fully developed Couette flow, using modified pressure. We are to compare to the solution of Chap. 9 that does not use modified pressure. Assumptions We number and list the assumptions for clarity: 1 The plates are infinite in x and z (z is out of the page in the figure associated with this problem). 2 The flow is steady. 3 This is a parallel flow (we assume the y component of velocity, v, is zero). 4 The fluid is incompressible and Newtonian, and the flow is laminar. 5 Pressure P = constant with respect to x. In other words, there is no applied pressure gradient pushing the flow in the x direction; the flow establishes itself due to viscous stresses caused by the moving upper wall. In terms of modified pressure, P′ is also constant with respect to x. 6 The velocity field is purely two-dimensional, which implies that w = 0 and ∂ ( any velocity component ) = 0 . ∂z 7 Gravity acts in the negative z direction. Analysis
To obtain the velocity and pressure fields, we follow the step-by-step procedure outlined in Chap. 9.
Step 1 Set up the problem and the geometry. See the figure associated with this problem. Step 2 List assumptions and boundary conditions. We have already listed seven assumptions. The boundary conditions come from imposing the no slip condition: (1) At the bottom plate (y = 0), u = v = w = 0. (2) At the top plate (y = h), u = V, v = 0, and w = 0. (3) At z = 0, P = P0, and thus P′ = P + ρgz = P0. Step 3 Write out and simplify the differential equations. We start with the continuity equation in Cartesian coordinates,
Continuity:
∂u + ∂x
∂v ∂y N
+
Assumption 3
∂w ∂z N
=0
∂u =0 ∂x
or
(1)
Assumption 6
Equation 1 tells us that u is not a function of x. In other words, it doesn’t matter where we place our origin – the flow is the same at any x location. I.e., the flow is fully developed. Furthermore, since u is not a function of time (Assumption 2) or z (Assumption 6), we conclude that u is at most a function of y, Result of continuity:
u = u ( y ) only
(2)
We now simplify the x momentum equation as far as possible: ⎛ ⎞ ⎛ ⎞ ⎜ ∂u ⎜ ∂ 2u ∂P ′ u u u ∂ 2u ∂ 2u ⎟ ∂ ∂ ∂ ⎟ ⎟= − ρ⎜ +μ⎜ + 2+ + u + v + w ⎟ ∂
∂t x ∂x 2 ∂y ∂z 2 ⎟ ∂x ∂y ∂z ⎟ ⎜ N ⎜ N N N N N Assumption 6 ⎟ ⎜ Continuity ⎟ ⎜ Assumption 2 Continuity Assumption Assumption 6 ⎠ ⎝ 3 ⎝ ⎠ Assumption 5
→
d 2u =0 dy 2
(3)
All other terms in Eq. 3 have disappeared except for a lone viscous term, which must then itself equal zero. Notice that we have changed from a partial derivative (∂/∂y) to a total derivative (d/dy) in Eq. 3 as a direct result of Eq. 2. We do not show the details here, but you can show in similar fashion that every term except the pressure term in the y momentum equation goes to zero, forcing that lone term to also be zero, ∂P ′ =0 ∂y
y momentum:
(4)
The same thing happens to the z momentum equation; the result is ∂P ′ =0 ∂z
z momentum:
(5)
In other words, P′ is not a function of y or z. Since P′ is also not a function of time (Assumption 2) or x (Assumption 5), P′ is a constant, Result of y and z momentum:
P ′ = constant = C3
(6)
10-7 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 10 Approximate Solutions of the Navier-Stokes Equation Step 4 Solve the differential equations. Continuity, y momentum, and z momentum have already been “solved”, resulting in Eqs. 2 and 6. Equation 3 (x momentum) is integrated twice to get u = C1 y + C2
Integration of x momentum:
(7)
where C1 and C2 are constants of integration. Step 5 We apply boundary condition (3), P′ = P0 at z = 0. Eq. 6 yields C3 = P0, and Final solution for pressure field:
P′ = P0
P = P0 − ρ gz
→
(9)
We next apply boundary conditions (1) and (2) to obtain constants C1 and C2. Boundary condition (1):
u = C1 (0) + C2 = 0
C2 = 0
or
and Boundary condition (2):
u = C1 (h) + 0 = V
or
C1 =
V h
Finally, Eq. 7 becomes
u =V
Final result for velocity field:
y h
(10)
The velocity field reveals a simple linear velocity profile from u = 0 at the bottom plate to u = V at the top plate. Step 6 Verify the results. You can plug in the velocity and pressure fields to verify that all the differential equations and boundary conditions are satisfied. We verify that the results are identical to those of Example 9-15. Thus, we get the same result using modified pressure throughout the calculation as we do using the regular (thermodynamic) pressure throughout the calculation. Discussion Since there are no free surfaces in this problem, the gravity term in the Navier-Stokes equation is absorbed into the modified pressure, and the pressure and gravity terms are combined into one term. This is possible since flow pressure and hydrostatic pressure are uncoupled.
10-8 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-13 Solution We are to write all three components of the Navier-Stokes equation in terms of modified pressure, and show that they are equivalent to the equations with regular pressure. We are also to discuss the advantage of using modified pressure. Analysis
In terms of modified pressure, the Navier-Stokes equation is written in Cartesian components as ⎛ ∂u ∂u ∂u ∂u ⎞ ∂P ′ +u +v +w ⎟=− + μ∇ 2 u ∂x ∂y ∂z ⎠ ∂x ⎝ ∂t
(1)
ρ⎜
⎛ ∂v ∂v ∂v ∂v ⎞ ∂P ′ +u +v + w ⎟ = − + μ∇ 2 v ∂x ∂y ∂z ⎠ ∂y ⎝ ∂t
(2)
⎛ ∂w ∂w ∂w ∂w ⎞ ∂P ′ +u +v +w + μ∇ 2 w ⎟=− t x y z z ∂ ∂ ∂ ∂ ∂ ⎝ ⎠
(3)
ρ⎜
x component: and
y component: and
ρ⎜
z component:
The definition of modified pressure is P ′ = P + ρ gz
Modified pressure:
(4)
When Eq. 4 is plugged into Eqs. 1 and 2, the gravity term disappears since z is independent of x and y. The result is ⎛ ∂u ∂u ∂u ∂u ⎞ ∂P +u +v +w ⎟=− + μ∇ 2 u ∂x ∂y ∂z ⎠ ∂x ⎝ ∂t
(5)
⎛ ∂v ∂v ∂v ∂v ⎞ ∂P +u +v + w ⎟ = − + μ∇ 2 v ∂ ∂ ∂ ∂ ∂ t x y z y ⎝ ⎠
(6)
ρ⎜
x component: and
ρ⎜
y component:
However, when Eq. 4 is plugged into Eq. 3, the result is
z component:
⎛ ∂w ∂w ∂w ∂w ⎞ ∂P +u +v +w − ρ g + μ∇ 2 w ⎟=− ∂ ∂ ∂ ∂ ∂ t x y z z ⎝ ⎠
ρ⎜
(7)
Equations 5 through 7 are the appropriate components of the Navier-Stokes equation in terms of regular pressure, so long as gravity acts downward (in the −z direction). The advantage of using modified pressure is that the gravity term disappears from the Navier-Stokes equation. Discussion
Modified pressure can be used only when there are no free surfaces.
10-9 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-14 Solution
We are to sketch the profile of actual pressure and shade in the region representing hydrostatic pressure.
Assumptions 1 The flow is incompressible. 2 Gravity acts vertically downward. 3 There are no free surface effects in this flow field.
P′ G g
Analysis By definition, modified pressure P′ = P + ρgz. Thus, to obtain actual pressure P, we subtract the hydrostatic component ρgz from the given profile of P′. Using the given value of P at the mid-way point as a guide, we sketch the actual pressure in Fig. 1 such that the difference between P′ and P increases linearly. In other words, we subtract the hydrostatic pressure component ρgz from the modified pressure P′ to obtain the profile for actual pressure P.
z
P
x
FIGURE 1 Actual pressure P (black arrows) and modified pressure P′ (gray arrows) for the given pressure field.
Discussion We assume that there are no free surface effects in the problem; otherwise modified pressure should not be used. The datum plane is set in the problem statement, but any arbitrary elevation could be used instead. If the datum plane were set at the top of the domain, P′ would be less than P everywhere because of the negative values of z in the transformation from P to P′.
10-15 Solution We are to solve the Navier-Stokes equation in terms of modified pressure for the case of steady, fully developed, laminar flow in a round pipe. We are to obtain expressions for the pressure and velocity fields, and compare the actual pressure at the top of the pipe to that at the bottom of the pipe. Assumptions pressure P.
We make the same assumptions as in Example 9-18, except we use modified pressure P′ in place of actual
Analysis The Navier-Stokes equation with gravity, written in terms of modified pressure P′, is identical to the Navier-Stokes equation with no gravity, written in terms of actual pressure P. In other words, all of the algebra of Example 9-18 remains the same, except we use modified pressure P′ in place of actual pressure P. The velocity field does not change, and the result is u=
Axial velocity field:
1 dP 2 r − R2 4 μ dx
(
)
(1)
The modified pressure field is P ′ = P ′ ( x ) = P1′ +
Modified pressure field:
dP ′ dx dx
(2)
where P1′ is the modified pressure at location x = x1. In Example 9-18, the actual pressure varies only with x. In fact it decreases linearly with x (note that the pressure gradient is negative for flow from left to right). Here, Eq. 2 shows that modified pressure behaves in the same way, namely P′ varies only with x, and in fact decreases linearly with x. We simply subtract the hydrostatic pressure component ρgz from modified pressure P′ (Eq. 2) to obtain the final expression for actual pressure P,
Actual pressure field:
P = P′ − ρ gz
→
P = P1′+
dP ′ dx − ρ gz dx
(3)
Since the pipe is horizontal, the bottom of the pipe is lower than the top of the pipe. Thus, ztop is greater than zbottom, and therefore by Eq. 3 Ptop is less than Pbottom. This agrees with our experience that pressure increases downward. Discussion Since there are no free surfaces in this flow, the gravity term does not directly influence the velocity field, and a hydrostatic component is added to the pressure field. You can see the advantage of using modified pressure.
10-10 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 10 Approximate Solutions of the Navier-Stokes Equation Creeping Flow
10-16C Solution We are to name each term in the Navier-Stokes equation, and then discuss which terms remain when the creeping flow approximation is made.
The terms in the equation are identified as follows:
Analysis
I II III IV V
Unsteady term Inertial term Pressure term Gravity term Viscous term
When the creeping flow approximation is made, only terms III (pressure) and V (viscous) remain. The other three terms are very small compared to these two and can be ignored. The significance is that all unsteady and inertial effects (terms I and II) have disappeared, as has gravity. We are left with a flow in which pressure forces and viscous forces must balance. Another significant result is that density has disappeared from the creeping flow equation, as discussed in the text. Discussion There are other acceptable one-word descriptions of some of the terms in the equation. For example, the inertial term can also be called the convective term, the advective term, or the acceleration term.
10-17 Solution We are to estimate the maximum speed of honey through a hole such that the Reynolds number remains below 0.1, at two different temperatures. Analysis
The density of honey is equal to its specific gravity times the density of water,
Density of honey:
ρ honey = SGhoney ρ water = 1.42 ( 998.0 kg/m3 ) = 1420 kg/m3
(1)
We convert the viscosity of honey from poise to standard SI units, Viscosity of honey at 20o: ⎛
⎞ ⎛ kg ⎞ ⎛ 100 cm ⎞ g kg = 19.0 ⎟⎜ ⎟⎜ ⎟ m ⋅s ⎝ cm ⋅ s ⋅ poise ⎠ ⎝ 1000 g ⎠ ⎝ m ⎠
μ honey = 190 poise ⎜
(2)
Finally, we plug Eqs. 1 and 2 into the definition of Reynolds number, and set Re = 0.1 to solve for the maximum speed to ensure creeping flow, Maximum speed for creeping flow at 20o: Re max μ honey ( 0.1)(19.0 kg/m ⋅ s ) Vmax = = = 0.33 m/s ρ honey D 1420 kg/m3 ( 0.0040 m )
(
)
(3)
At the higher temperature of 40oC, the calculations yield Vmax = 0.035 m/s. Thus, it is much easier to achieve creeping flow with honey at lower temperatures since the viscosity of honey increases rapidly as the temperature drops. Discussion We used Re < 0.1 as the maximum Reynolds number for creeping flow, but experiments reveal that in many flows, the creeping flow approximation is acceptable at Reynolds numbers as high as nearly 1.0.
10-11 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 10 Approximate Solutions of the Navier-Stokes Equation 10-18 Solution appropriate.
For each case we are to calculate the Reynolds number and determine if the creeping flow approximation is
Assumptions
1 The values given are characteristic scales of the motion.
Properties For water at T = 20oC, ρ = 998.0 kg/m3 and μ = 1.002 × 10-3 kg/m⋅s. For unused engine oil at T = 140oC, ρ 3 = 816.8 kg/m and μ = 6.558 × 10-3 kg/m⋅s. For air at T = 30oC, ρ = 1.164 kg/m3 and μ = 1.872 × 10-5 kg/m⋅s. Analysis
(a) The Reynolds number of the microorganism is Re =
3 −6 ρ DV ( 998.0 kg/m )( 5.0 × 10 m ) ( 0.2 mm/s ) ⎛ m ⎞ -4 = ⎜ 1000 mm ⎟ = 9.96 ×10 μ 1.002 × 10 −3 kg/m ⋅ s ⎝ ⎠
Since Re 1, the creeping flow approximation is not appropriate. (c) The Reynolds number of the fog droplet is Re =
3 −6 m ρ DV (1.164 kg/m )(10 × 10 m ) ( 3.0 mm/s ) ⎛ ⎞ -3 = ⎜ 1000 mm ⎟ = 1.87 ×10 1.872 × 10−5 kg/m ⋅ s μ ⎝ ⎠
Since Re 1) to subsonic (Ma < 1).
12-74C Solution
We are to analyze a claim about oblique shock analysis.
Analysis Yes, the claim is correct. Conversely, normal shocks can be thought of as special oblique shocks in which the shock angle is β = π/2, or 90o. Discussion The component of flow in the direction normal to the oblique shock acts exactly like a normal shock. We can think of the flow parallel to the oblique shock as “going along for the ride” – it does not affect anything.
12-31 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-75C Solution
We are to discuss shock detachment at the nose of a 2-D wedge-shaped body.
Analysis When the wedge half-angle δ is greater than the maximum deflection angle θmax, the shock becomes curved and detaches from the nose of the wedge, forming what is called a detached oblique shock or a bow wave. The numerical value of the shock angle at the nose is β = 90o. Discussion
When δ is less than θmax, the oblique shock is attached to the nose.
12-76C Solution
We are to discuss the shock at the nose of a rounded body in supersonic flow.
Analysis When supersonic flow impinges on a blunt body like the rounded nose of an aircraft, the wedge half-angle δ at the nose is 90o, and an attached oblique shock cannot exist, regardless of Mach number. Therefore, a detached oblique shock must occur in front of all such blunt-nosed bodies, whether two-dimensional, axisymmetric, or fully threedimensional. Discussion
Since δ = 90o at the nose, δ is always greater than θmax, regardless of Ma or the shape of the rest of the body.
12-77C Solution
We are to discuss the applicability of the isentropic flow relations across shocks and expansion waves.
Analysis The isentropic relations of ideal gases are not applicable for flows across (a) normal shock waves and (b) oblique shock waves, but they are applicable for flows across (c) Prandtl-Meyer expansion waves. Discussion
Flow across any kind of shock wave involves irreversible losses – hence, it cannot be isentropic.
12-78 Solution developed.
For an ideal gas flowing through a normal shock, a relation for V2/V1 in terms of k, Ma1, and Ma2 is to be
Analysis
The conservation of mass relation across the shock is ρ1V1 = ρ 2V2 and it can be expressed as V2 ρ1 P / RT1 ⎛ P1 = = 1 =⎜ V1 ρ 2 P2 / RT2 ⎜⎝ P2
⎞⎛ T2 ⎟⎟⎜⎜ ⎠⎝ T1
⎞ ⎟⎟ ⎠
From Eqs. 12-35 and 12-38, V 2 ⎛ 1 + kMa 22 =⎜ V1 ⎜⎝ 1 + kMa 12
⎞⎛ 1 + Ma 12 (k − 1) / 2 ⎞ ⎟⎜ ⎟ ⎟⎜ 1 + Ma 2 (k − 1) / 2 ⎟ 2 ⎠⎝ ⎠
Discussion This is an important relation as it enables us to determine the velocity ratio across a normal shock when the Mach numbers before and after the shock are known.
12-32 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-79 Solution Air flowing through a converging-diverging nozzle experiences a normal shock at the exit. The effect of the shock wave on various properties is to be determined. Assumptions 1 Air is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic before the shock occurs. 3 The shock wave occurs at the exit plane. Properties
The properties of air are k = 1.4 and R = 0.287 kJ/kg·K.
Analysis The inlet stagnation properties in this case are identical to the inlet properties since the inlet velocity is negligible. Then,
P01 = Pi = 1 MPa T01 = Ti = 300 K
Shock wave
Then, ⎛ 2 T1 = T01 ⎜⎜ 2 ( 1)Ma 12 + − k ⎝
⎞ ⎛ 2 ⎟ = (300 K)⎜ ⎜ 2 + (1.4 - 1)2 2 ⎟ ⎝ ⎠
⎞ ⎟ = 166.7 K ⎟ ⎠
AIR
i
1
2
Vi ≈ 0
and ⎛T P1 = P01 ⎜⎜ 1 ⎝ T0
⎞ ⎟ ⎟ ⎠
k /( k −1)
⎛ 166.7 ⎞ = (1 MPa)⎜ ⎟ ⎝ 300 ⎠
1.4 / 0.4
= 0.1278 MPa
The fluid properties after the shock (denoted by subscript 2) are related to those before the shock through the functions listed in Table A-14. For Ma1 = 2.0 we read Ma 2 = 0.5774 ,
P02 P T = 0.7209, 2 = 4.5000, and 2 = 1.6875 P02 P1 T1
Then the stagnation pressure P02, static pressure P2, and static temperature T2, are determined to be
P02 = 0.7209P01 = (0.7209)(1.0 MPa) = 0.721 MPa P2 = 4.5000P1 = (4.5000)(0.1278 MPa) = 0.575 MPa T2 = 1.6875T1 = (1.6875)(166.7 K) = 281 K The air velocity after the shock can be determined from V2 = Ma2c2, where c2 is the speed of sound at the exit conditions after the shock, ⎛ 1000 m 2 / s 2 V2 = Ma2c2 = Ma 2 kRT2 = (0.5774) (1.4)(0.287 kJ/kg ⋅ K)(281 K)⎜ ⎜ 1 kJ/kg ⎝
Discussion identical.
⎞ ⎟ = 194 m/s ⎟ ⎠
We can also solve this problem using the relations for normal shock functions. The results would be
12-33 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-80 Solution Air enters a converging-diverging nozzle at a specified state. The required back pressure that produces a normal shock at the exit plane is to be determined for the specified nozzle geometry. Assumptions 1 Air is an ideal gas. 2 Flow through the nozzle is steady, one-dimensional, and isentropic before the shock occurs. 3 The shock wave occurs at the exit plane. Analysis The inlet stagnation pressure in this case is identical to the inlet pressure since the inlet velocity is negligible. Since the flow before the shock to be isentropic,
P01 = Pi = 2 MPa
shock wave
It is specified that A/A* =3.5. From Table A-13, Mach number and the pressure ratio which corresponds to this area ratio are the Ma1 =2.80 and P1/P01 = 0.0368. The pressure ratio across the shock for this Ma1 value is, from Table A-14, P2/P1 = 8.98. Thus the back pressure, which is equal to the static pressure at the nozzle exit, must be
i
AIR
2
1
Vi ≈ 0 Pb
P2 =8.98P1 = 8.98×0.0368P01 = 8.98×0.0368×(2 MPa) = 0.661 MPa We can also solve this problem using the relations for compressible flow and normal shock functions. The Discussion results would be identical.
12-81 Solution Air enters a converging-diverging nozzle at a specified state. The required back pressure that produces a normal shock at the exit plane is to be determined for the specified nozzle geometry. shock wave
Assumptions 1 Air is an ideal gas. 2 Flow through the nozzle is steady, one-dimensional, and isentropic before the shock occurs. Analysis The inlet stagnation pressure in this case is identical to the inlet pressure since the inlet velocity is negligible. Since the flow before the shock to be isentropic,
i
AIR
1
2
Vi ≈ 0 Pb
P0x= Pi = 2 MPa It is specified that A/A* = 2. From Table A-13, the Mach number and the pressure ratio which corresponds to this area ratio are the Ma1 =2.20 and P1/P01 = 0.0935. The pressure ratio across the shock for this M1 value is, from Table A-14, P2/P1 = 5.48. Thus the back pressure, which is equal to the static pressure at the nozzle exit, must be
P2 =5.48P1 = 5.48×0.0935P01 = 5.48×0.0935×(2 MPa) = 1.02 MPa Discussion We can also solve this problem using the relations for compressible flow and normal shock functions. The results would be identical.
12-34 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-82 Solution Air flowing through a nozzle experiences a normal shock. The effect of the shock wave on various properties is to be determined. Analysis is to be repeated for helium under the same conditions. Assumptions 1 Air and helium are ideal gases with constant specific heats. 2 Flow through the nozzle is steady, onedimensional, and isentropic before the shock occurs. Properties The properties of air are k = 1.4 and R = 0.287 kJ/kg·K, and the properties of helium are k = 1.667 and R = 2.0769 kJ/kg·K. shock
The air properties upstream the shock are wave Ma1 = 2.5, P1 = 61.64 kPa, and T1 = 262.15 K AIR Fluid properties after the shock (denoted by subscript 2) are related to those i 2 1 before the shock through the functions in Table A-14. For Ma1 = 2.5, Ma1 = 2.5 P P T Ma 2 = 0.513, 02 = 8.5262, 2 = 7.125, and 2 = 2.1375 P1 P1 T1 Then the stagnation pressure P02, static pressure P2, and static temperature T2, are determined to be P02 = 8.5261P1 = (8.5261)(61.64 kPa) = 526 kPa P2 = 7.125P1 = (7.125)(61.64 kPa) = 439 kPa T2 = 2.1375T1 = (2.1375)(262.15 K) = 560 K The air velocity after the shock can be determined from V2 = Ma2c2, where c2 is the speed of sound at the exit conditions after the shock, Analysis
⎛ 1000 m 2 / s 2 ⎞ ⎟ = 243 m/s V 2 = Ma 2 c 2 = Ma 2 kRT2 = (0.513) (1.4)(0.287 kJ/kg ⋅ K)(560.3 K)⎜⎜ ⎟ ⎝ 1 kJ/kg ⎠ We now repeat the analysis for helium. This time we cannot use the tabulated values in Table A-14 since k is not 1.4. Therefore, we have to calculate the desired quantities using the analytical relations, ⎛ Ma 12 + 2 /(k − 1) ⎞ ⎟ Ma 2 = ⎜ ⎜ 2Ma 2 k /( k − 1) − 1 ⎟ 1 ⎝ ⎠
1/ 2
⎛ ⎞ 2.5 2 + 2 /(1.667 − 1) ⎟ =⎜ ⎜ 2 × 2.5 2 × 1.667 /(1.667 − 1) − 1 ⎟ ⎝ ⎠
1/ 2
= 0.553
P2 1 + kMa 12 1 + 1.667 × 2.5 2 = = = 7.5632 P1 1 + kMa 22 1 + 1.667 × 0.553 2 T2 1 + Ma 12 (k − 1) / 2 1 + 2.5 2 (1.667 − 1) / 2 = = = 2.7989 T1 1 + Ma 22 (k − 1) / 2 1 + 0.553 2 (1.667 − 1) / 2 P02 ⎛ 1 + kMa 12 =⎜ P1 ⎜⎝ 1 + kMa 22
⎞ ⎟ 1 + (k − 1)Ma 22 / 2 ⎟ ⎠
(
⎛ 1 + 1.667 × 2.5 2 = ⎜⎜ 2 ⎝ 1 + 1.667 × 0.553 Thus,
)k /(k −1)
⎞ 1.667 / 0.667 ⎟ 1 + (1.667 − 1) × 0.553 2 / 2 = 9.641 ⎟ ⎠
(
)
P02 = 11.546P1 = (11.546)(61.64 kPa) = 712 kPa P2 = 7.5632P1 = (7.5632)(61.64 kPa) = 466 kPa T2 = 2.7989T1 = (2.7989)(262.15 K) = 734 K
⎛ 1000 m 2 / s 2 ⎞ ⎟ = 881 m/s V 2 = Ma 2 c 2 = Ma 2 kRT y = (0.553) (1.667)(2.0769 kJ/kg ⋅ K)(733.7 K)⎜⎜ ⎟ ⎝ 1 kJ/kg ⎠ Discussion The velocity and Mach number are higher for helium than for air due to the different values of k and R.
12-35 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-83 Solution Air flowing through a nozzle experiences a normal shock. The entropy change of air across the normal shock wave is to be determined. Assumptions 1 Air and helium are ideal gases with constant specific heats. 2 Flow through the nozzle is steady, onedimensional, and isentropic before the shock occurs. Properties The properties of air are R = 0.287 kJ/kg·K and cp = 1.005 kJ/kg·K, and the properties of helium are R = 2.0769 kJ/kg·K and cp = 5.1926 kJ/kg·K. Analysis
The entropy change across the shock is determined to be
s 2 − s1 = c p ln
T2 P − R ln 2 = (1.005 kJ/kg ⋅ K)ln(2.1375) - (0.287 kJ/kg ⋅ K)ln(7.125) = 0.200 kJ/kg ⋅ K T1 P1
For helium, the entropy change across the shock is determined to be s 2 − s1 = c p ln
T2 P − R ln 2 = (5.1926 kJ/kg ⋅ K)ln(2.7989) - (2.0769 kJ/kg ⋅ K)ln(7.5632) = 1.14 kJ/kg ⋅ K T1 P1
Discussion waves.
Note that shock wave is a highly dissipative process, and the entropy generation is large during shock
12-36 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-84E [Also solved using EES on enclosed DVD]
Solution Air flowing through a nozzle experiences a normal shock. Effect of the shock wave on various properties is to be determined. Analysis is to be repeated for helium Assumptions 1 Air and helium are ideal gases with constant specific heats. 2 Flow through the nozzle is steady, onedimensional, and isentropic before the shock occurs. Properties The properties of air are k = 1.4 and R = 0.06855 Btu/lbm·R, and the properties of helium are k = 1.667 and R = 0.4961 Btu/lbm·R. Analysis
shock wave
The air properties upstream the shock are Ma1 = 2.5, P1 = 10 psia, and T1 = 440.5 R i
Fluid properties after the shock (denoted by subscript 2) are related to those before the shock through the functions listed in Table A-14. For Ma1 = 2.5, P P T Ma 2 = 0.513, 02 = 8.5262, 2 = 7.125, and 2 = 2.1375 P1 P1 T1
AIR
2
1
Ma1 = 2.5
Then the stagnation pressure P02, static pressure P2, and static temperature T2, are determined to be P02 = 8.5262P1 = (8.5262)(10 psia) = 85.3 psia P2 = 7.125P1 = (7.125)(10 psia) = 71.3 psia T2 = 2.1375T1 = (2.1375)(440.5 R) = 942 R The air velocity after the shock can be determined from V2 = Ma2c2, where c2 is the speed of sound at the exit conditions after the shock, ⎛ 25,037 ft 2 / s 2 V 2 = Ma 2 c 2 = Ma 2 kRT2 = (0.513) (1.4)(0.06855 Btu/1bm ⋅ R)(941.6 R)⎜ ⎜ 1 Btu/1bm ⎝
⎞ ⎟ = 772 ft/s ⎟ ⎠
We now repeat the analysis for helium. This time we cannot use the tabulated values in Table A-14 since k is not 1.4. Therefore, we have to calculate the desired quantities using the analytical relations, ⎛ Ma 12 + 2 /(k − 1) ⎞ ⎟ Ma 2 = ⎜ ⎜ 2Ma 2 k /( k − 1) − 1 ⎟ 1 ⎝ ⎠
1/ 2
⎛ ⎞ 2.5 2 + 2 /(1.667 − 1) ⎟ =⎜ ⎜ 2 × 2.5 2 × 1.667 /(1.667 − 1) − 1 ⎟ ⎝ ⎠
1/ 2
= 0.553
P2 1 + kMa 12 1 + 1.667 × 2.5 2 = = = 7.5632 P1 1 + kMa 22 1 + 1.667 × 0.553 2 T2 1 + Ma 12 (k − 1) / 2 1 + 2.5 2 (1.667 − 1) / 2 = = = 2.7989 T1 1 + Ma 22 (k − 1) / 2 1 + 0.553 2 (1.667 − 1) / 2 P02 ⎛ 1 + kMa 12 =⎜ P1 ⎜⎝ 1 + kMa 22
⎞ ⎟ 1 + (k − 1)Ma 22 / 2 ⎟ ⎠
(
⎛ 1 + 1.667 × 2.5 2 = ⎜⎜ 2 ⎝ 1 + 1.667 × 0.553 Thus,
)k /(k −1)
⎞ 1.667 / 0.667 ⎟ 1 + (1.667 − 1) × 0.553 2 / 2 = 9.641 ⎟ ⎠
(
)
P02 = 11.546P1 = (11.546)(10 psia) = 115 psia P2 = 7.5632P1 = (7.5632)(10 psia) = 75.6 psia T2 = 2.7989T1 = (2.7989)(440.5 R) = 1233 R
⎛ 25,037 ft 2 / s 2 ⎞ ⎟ = 2794 ft/s V 2 = Ma 2 c 2 = Ma 2 kRT2 = (0.553) (1.667)(0.4961 Btu/1bm.R)(1232.9 R)⎜⎜ ⎟ ⎝ 1 Btu/1bm ⎠ Discussion This problem could also be solved using the relations for compressible flow and normal shock functions. The results would be identical.
12-37 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-85E
Solution We are to reconsider Prob. 12-84E. Using EES (or other) software, we are to study the effects of both air and helium flowing steadily in a nozzle when there is a normal shock at a Mach number in the range 2 < Mx < 3.5. In addition to the required information, we are to calculate the entropy change of the air and helium across the normal shock, and tabulate the results in a parametric table.
We use EES to calculate the entropy change of the air and helium across the normal shock. The results are Analysis given in the Parametric Table for 2 < M_x < 3.5. Procedure NormalShock(M_x,k:M_y,PyOPx, TyOTx,RhoyORhox, PoyOPox, PoyOPx) If M_x < 1 Then M_y = -1000;PyOPx=-1000;TyOTx=-1000;RhoyORhox=-1000 PoyOPox=-1000;PoyOPx=-1000 else M_y=sqrt( (M_x^2+2/(k-1)) / (2*M_x^2*k/(k-1)-1) ) PyOPx=(1+k*M_x^2)/(1+k*M_y^2) TyOTx=( 1+M_x^2*(k-1)/2 )/(1+M_y^2*(k-1)/2 ) RhoyORhox=PyOPx/TyOTx PoyOPox=M_x/M_y*( (1+M_y^2*(k-1)/2)/ (1+M_x^2*(k-1)/2) )^((k+1)/(2*(k-1))) PoyOPx=(1+k*M_x^2)*(1+M_y^2*(k-1)/2)^(k/(k-1))/(1+k*M_y^2) Endif End Function ExitPress(P_back,P_crit) If P_back>=P_crit then ExitPress:=P_back If P_back 94 kPa
Therefore, the flow is choked, and the velocity at the exit of the hole is the sonic speed. Then the flow properties at the exit becomes P 314 kPa = 3.671 kg/m 3 ρ0 = 0 = RT0 (0.287 kPa ⋅ m 3 / kg ⋅ K )(298 K)
ρ * = ρ 0 ⎛⎜
2 ⎞ ⎟ k ⎝ + 1⎠
T* =
1 /( k −1)
2 ⎞ = (3.671 kg/m 3 )⎛⎜ ⎟ + 1⎠ 1 . 4 ⎝
1 /(1.4 −1)
= 2.327 kg/m 3
2 2 T0 = (298 K) = 248.3 K k +1 1.4 + 1
⎛ 1000 m 2 / s 2 V = c = kRT * = (1.4 )(0.287 kJ/kg ⋅ K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟(248.3 K) = 315.9 m/s ⎟ ⎠
Then the initial mass flow rate through the hole becomes m = ρAV = (2.327 kg/m 3 )[π (0.004 m) 2 /4](315.9 m/s) = 0.00924 kg/s = 0.554 kg/min
Discussion
The mass flow rate will decrease with time as the pressure inside the tire drops.
12-73 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-132 Solution The thrust developed by the engine of a Boeing 777 is about 380 kN. The mass flow rate of gases through the nozzle is to be determined. Assumptions 1 Air is an ideal gas with constant specific heats. 2 Flow of combustion gases through the nozzle is isentropic. 3 Choked flow conditions exist at the nozzle exit. 4 The velocity of gases at the nozzle inlet is negligible. Properties The gas constant of air is R = 0.287 kPa⋅m3/kg⋅K, and it can also be used for combustion gases. The specific heat ratio of combustion gases is k = 1.33.
The velocity at the nozzle exit is the sonic speed, which is determined to be
Analysis
⎛ 1000 m 2 / s 2 V = c = kRT = (1.33)(0.287 kJ/kg ⋅ K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟(295 K) = 335.6 m/s ⎟ ⎠
Noting that thrust F is related to velocity by F = m V , the mass flow rate of combustion gases is determined to be m =
F 380 , 000 N ⎛ 1 kg.m/s 2 = ⎜ V 335.6 m/s ⎝ 1 N
⎞ ⎟ = 1132 kg/s ≅ 1130 kg/s ⎠
Discussion The combustion gases are mostly nitrogen (due to the 78% of N2 in air), and thus they can be treated as air with a good degree of approximation.
12-133 Solution determined.
A stationary temperature probe is inserted into an air duct reads 85°C. The actual temperature of air is to be
Assumptions
1 Air is an ideal gas with constant specific heats at room temperature. 2 The stagnation process is isentropic.
Properties
The specific heat of air at room temperature is cp = 1.005 kJ/kg⋅K.
Analysis The air that strikes the probe will be brought to a complete stop, and thus it will undergo a stagnation process. The thermometer will sense the temperature of this stagnated air, which is the stagnation temperature. The actual air temperature is determined from T = T0 −
(250 m/s) 2 V2 ⎛ 1 kJ/kg ⎞ = 85°C − ⎜ ⎟ = 53.9°C 2c p 2 × 1.005 kJ/kg ⋅ K ⎝ 1000 m 2 / s 2 ⎠
T
250 m/s
Discussion Temperature rise due to stagnation is very significant in high-speed flows, and should always be considered when compressibility effects are not negligible.
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Chapter 12 Compressible Flow 12-134 Solution Nitrogen flows through a heat exchanger. The stagnation pressure and temperature of the nitrogen at the inlet and the exit states are to be determined. Assumptions isentropic.
1 Nitrogen is an ideal gas with constant specific heats. 2 Flow of nitrogen through the heat exchanger is
Properties
The properties of nitrogen are cp = 1.039 kJ/kg⋅K and k = 1.4.
Analysis
The stagnation temperature and pressure of nitrogen at the inlet and the exit states are determined from
T01 = T1 +
V1 2 (100 m/s) 2 ⎛ 1 kJ/kg ⎞ = 10°C + ⎜ ⎟ = 14.8°C 2c p 2 × 1.039 kJ/kg ⋅ °C ⎝ 1000 m 2 / s 2 ⎠
⎛T ⎞ P01 = P1 ⎜⎜ 01 ⎟⎟ ⎝ T1 ⎠
k /( k −1)
⎛ 288.0 K ⎞ = (150 kPa)⎜ ⎟ ⎝ 283.2 K ⎠
1.4 /(1.4 −1)
= 159 kPa
Qin
150 kPa 10°C 100 m/s
Nitrogen
100 kPa 200 m/s
From the energy balance relation E in − E out = ΔE system with w = 0 V 22 − V1 2 + Δpe 0 2 (200 m/s) 2 − (100 m/s) 2 ⎛ 1 kJ/kg ⎞ 150 kJ/kg = (1.039 kJ/kg ⋅ °C)(T2 − 10°C) + ⎜ ⎟ 2 ⎝ 1000 m 2 / s 2 ⎠ T2 = 139.9°C q in = c p (T2 − T1 ) +
and T02 = T2 +
P02
V2 2 (200 m/s) 2 ⎛ 1 kJ/kg ⎞ = 139.9°C + ⎜ ⎟ = 159°C 2 × 1.039 kJ/kg ⋅ °C ⎝ 1000 m 2 / s 2 ⎠ 2c p
⎛T = P2 ⎜⎜ 02 ⎝ T2
⎞ ⎟⎟ ⎠
k /( k −1)
⎛ 432.3 K ⎞ = (100 kPa)⎜ ⎟ ⎝ 413.1 K ⎠
1.4 /(1.4 −1)
= 117 kPa
Discussion Note that the stagnation temperature and pressure can be very different than their thermodynamic counterparts when dealing with compressible flow.
12-75 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-135 Solution An expression for the speed of sound based on van der Waals equation of state is to be derived. Using this relation, the speed of sound in carbon dioxide is to be determined and compared to that obtained by ideal gas behavior. Properties
The properties of CO2 are R = 0.1889 kJ/kg·K and k = 1.279 at T = 50°C = 323.2 K.
Analysis
Van der Waals equation of state can be expressed as P =
Differentiating,
RT a − 2 . v−b v
⎛ ∂P ⎞ = RT + 2a ⎜ ⎟ ⎝ ∂v ⎠ T (v − b) 2 v 3
Noting that ρ = 1 / v ⎯ ⎯→ dρ = −dv / v 2 , the speed of sound relation becomes
Substituting,
⎛ ∂P ⎞ ⎛ ∂P ⎞ c 2 = k⎜ ⎟ = v 2 k⎜ ⎟ ∂ r ⎝ ⎠T ⎝ ∂v ⎠ T c2 =
v 2 kRT (v − b ) 2
−
2ak v
Using the molar mass of CO2 (M = 44 kg/kmol), the constant a and b can be expressed per unit mass as a = 0.1882 kPa ⋅ m 6 /kg 2 and
b = 9.70 × 10 −4 m 3 / kg
The specific volume of CO2 is determined to be 200 kPa =
(0.1889 kPa ⋅ m 3 / kg ⋅ K)(323.2 K) v − 0.000970 m 3 /kg
−
2 × 0.1882 kPa ⋅ m 6 / kg 2 v2
→ v = 0.300 m 3 / kg
Substituting, ⎛ ⎛ (0.300 m 3 / kg) 2 (1.279)(0.1889 kJ/kg ⋅ K)(323.2 K) 2(0.1882 kPa ⋅ m 6 / kg 3 )(1.279) ⎞⎛ 1000 m 2 / s 2 ⎞ ⎞ ⎟⎜ ⎟⎟ − c = ⎜⎜ 3 2 3 2 ⎟⎜ 1 kPa ⋅ m 3 /kg ⎟ ⎟ ⎜⎜ − (0.300 0 . 0 00970 m / kg ) ( 0 . 300 m / kg) ⎠⎝ ⎠⎠ ⎝⎝ = 271 m/s
1/ 2
If we treat CO2 as an ideal gas, the speed of sound becomes ⎛ 1000 m 2 / s 2 c = kRT = (1.279)(0.1889 kJ/kg ⋅ K)(323.2 K)⎜ ⎜ 1 kJ/kg ⎝
⎞ ⎟ = 279 m/s ⎟ ⎠
Discussion Note that the ideal gas relation is the simplest equation of state, and it is very accurate for most gases encountered in practice. At high pressures and/or low temperatures, however, the gases deviate from ideal gas behavior, and it becomes necessary to use more complicated equations of state.
12-76 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-136 Solution
The equivalent relation for the speed of sound is to be verified using thermodynamic relations. ⎛ ∂P ⎞ ⎛ ∂P ⎞ The two relations are c 2 = ⎜⎜ ⎟⎟ and c 2 = k ⎜⎜ ⎟⎟ ⎝ ∂ρ ⎠ s ⎝ ∂ρ ⎠ T
Analysis
From r = 1 / v
∂P ∂P ∂P ∂T ⎞ 2 ⎛ ∂P ⎞ ⎛ ∂T ⎞ dr = − dv / v 2 . Thus, c 2 = ⎛⎜ ⎞⎟ = −v 2 ⎛⎜ ⎞⎟ = −v 2 ⎛⎜ ⎟ = −v ⎜ ⎟ ⎜ ⎟ ⎝ ∂r ⎠ s ⎝ ∂v ⎠ s ⎝ ∂T ∂v ⎠ s ⎝ ∂T ⎠ s ⎝ ∂v ⎠ s
⎯ ⎯→
From the cyclic rule,
∂P ⎞ ⎛ ∂T ⎞ ⎛ ∂s ⎞ ∂P ⎞ ⎛ ∂s ⎞ ⎛ ∂P ⎞ ( P, T , s ) : ⎛⎜ ⎯→ ⎛⎜ ⎟ ⎜ ⎟ ⎜ ⎟ = −1 ⎯ ⎟ = −⎜ ⎟ ⎜ ⎟ ∂ T ∂ s ∂ P ∂ T ⎝ ⎠s ⎝ ⎠ P ⎝ ⎠T ⎝ ⎠s ⎝ ∂T ⎠ P ⎝ ∂s ⎠ T
⎛∂ s ⎞ ⎛∂T ⎞ ⎛∂T ⎞ ⎛∂T ⎞ ⎛∂v⎞ ⎛ ∂ s ⎞ ⎟⎟ ⎟⎟ = −⎜⎜ ⎟⎟ ⎜⎜ ⎟⎟ = −1 ⎯ ⎟⎟ ⎜⎜ ⎟⎟ ⎜⎜ (T , v, s ) : ⎜⎜ ⎯→ ⎜⎜ ⎝ ∂ v ⎠T ⎝ ∂ s ⎠ v ⎝ ∂ v ⎠s ⎝ ∂ v ⎠ s ⎝ ∂ s ⎠T ⎝ ∂ T ⎠ v Substituting, ⎛ ∂s ⎞ ⎛ ∂P ⎞ ⎛ ∂s ⎞ ⎛ ∂T ⎞ 2 ⎛ ∂s ⎞ ⎛ ∂T ⎞ ⎛ ∂P ⎞ c 2 = −v 2 ⎜ ⎟ ⎜ ⎟ ⎜ ⎟ ⎜ ⎟ = −v ⎜ ⎟ ⎜ ⎟ ⎜ ⎟ ⎝ ∂T ⎠ P ⎝ ∂s ⎠ T ⎝ ∂v ⎠ T ⎝ ∂s ⎠ v ⎝ ∂T ⎠ P ⎝ ∂s ⎠ v ⎝ ∂v ⎠ T
Recall that
⎛ ∂s ⎞ =⎜ ⎟ and T ⎝ ∂T ⎠ P
cp
⎛ cp c 2 = −v 2 ⎜⎜ ⎝ T
⎞⎛ T ⎟⎟⎜⎜ ⎠⎝ c v
c v ⎛ ∂s ⎞ =⎜ ⎟ . Substituting, T ⎝ ∂T ⎠ v
⎞⎛ ∂P ⎞ ∂P ⎟⎜ ⎟ = −v 2 k ⎛⎜ ⎞⎟ ⎟⎝ ∂v ⎠ ∂v ⎠ T ⎝ T ⎠
⎛ ∂P ⎞ Replacing − dv / v 2 by dρ, we get c 2 = k ⎜⎜ ⎟⎟ , which is the desired expression ⎝ ∂ρ ⎠ T
Discussion Note that the differential thermodynamic property relations are very useful in the derivation of other property relations in differential form.
12-137 Solution For ideal gases undergoing isentropic flows, expressions for P/P*, T/T*, and ρ/ρ* as functions of k and Ma are to be obtained.
Equations 12-18 and 12-21 are given to be
Analysis
Multiplying the two,
⎛ T0 T * ⎞ ⎛ 2 + (k − 1)Ma 2 ⎟ ⎜ ⎜ ⎜ T T ⎟=⎜ 2 0 ⎠ ⎝ ⎝
Simplifying and inverting,
T k +1 = T * 2 + (k − 1)Ma 2
P ⎛ T ⎞ =⎜ ⎟ P* ⎝T *⎠
k /( k −1)
From
ρ ⎛ ρ ⎞ ⎟ =⎜ ρ * ⎜⎝ ρ * ⎟⎠
k /( k −1)
From
Discussion
T0 2 + (k − 1)Ma 2 = T 2
and
2 T* = T0 k + 1
⎞⎛ 2 ⎞ ⎟⎜ ⎟⎝ k + 1 ⎟⎠ ⎠
(1)
P ⎛ k +1 =⎜ P * ⎜⎝ 2 + (k − 1)Ma 2
⎞ ⎟ ⎟ ⎠
k /( k −1)
⎯ ⎯→
ρ ⎛⎜ k +1 = ρ * ⎜⎝ 2 + (k − 1)Ma 2
⎞ ⎟ ⎟ ⎠
k /( k −1)
⎯ ⎯→
(2)
(3)
Note that some very useful relations can be obtained by very simple manipulations.
12-77 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-138 Solution It is to be verified that for the steady flow of ideal gases dT0/T = dA/A + (1-Ma2) dV/V. The effect of heating and area changes on the velocity of an ideal gas in steady flow for subsonic flow and supersonic flow are to be explained.
V2 = c p (T0 − T ) 2 V dV = c p (dT0 − dT )
(1)
We start with the relation
Analysis
Differentiating, We also have
dρ
ρ dP
and
ρ
+
(2)
dA dV + =0 A V
(3)
+ V dV = 0
(4) dP dρ dT = + =0 ρ P T
Differentiating the ideal gas relation P = ρRT,
(5)
From the speed of sound relation,
c 2 = kRT = (k − 1)c p T = kP / ρ
(6)
Combining Eqs. (3) and (5),
dP dT dA dV − + + =0 P T A V
(7)
Combining Eqs. (4) and (6),
dP
ρ
=
dP kP / c 2
= −VdV
or,
V 2 dV dV dP k = − 2 V dV = − k 2 = −kMa 2 V P C V C
Combining Eqs. (2) and (6),
dT = dT0 − V
or,
(8)
dV cp
dV dT dT0 V 2 dV dT dT0 V2 dV dT0 = − = = − 2 = − (k − 1)Ma 2 T T C pT V T T T V C /( k − 1) V
(9)
dV dT0 dV dA dV − + (k − 1)Ma 2 + + =0 V T V A V
Combining Eqs. (7), (8), and (9),
− ( k − 1)Ma 2
or,
dT0 dA dV = + − kMa 2 + (k − 1)Ma 2 + 1 T A V
Thus,
dT0 dA dV = + (1 − Ma 2 ) T A V
[
]
(10)
Differentiating the steady-flow energy equation q = h02 − h01 = c p (T02 − T01 )
δq = c p dT0
(11)
Eq. (11) relates the stagnation temperature change dT0 to the net heat transferred to the fluid. Eq. (10) relates the velocity changes to area changes dA, and the stagnation temperature change dT0 or the heat transferred. (a) When Ma < 1 (subsonic flow), the fluid accelerates if the duct converges (dA < 0) or the fluid is heated (dT0 > 0 or δq > 0). The fluid decelerates if the duct converges (dA < 0) or the fluid is cooled (dT0 < 0 or δq < 0). (b) When Ma > 1 (supersonic flow), the fluid accelerates if the duct diverges (dA > 0) or the fluid is cooled (dT0 < 0 or δq < 0). The fluid decelerates if the duct converges (dA < 0) or the fluid is heated (dT0 > 0 or δq > 0). Discussion
Some of these results are not intuitively obvious, but come about by satisfying the conservation equations.
12-78 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-139 Solution A Pitot-static probe measures the difference between the static and stagnation pressures for a subsonic airplane. The speed of the airplane and the flight Mach number are to be determined. Assumptions
1 Air is an ideal gas with a constant specific heat ratio. 2 The stagnation process is isentropic.
Properties
The properties of air are R = 0.287 kJ/kg⋅K and k = 1.4.
Analysis
The stagnation pressure of air at the specified conditions is P0 = P + ΔP = 70.109 + 22 = 92.109 kPa
Then, P0 ⎛ (k − 1)Ma 2 = ⎜1 + P ⎜⎝ 2
⎞ ⎟ ⎟ ⎠
k / k −1
⎯ ⎯→
92.109 ⎛ (1.4 − 1)Ma 2 = ⎜1 + 70.109 ⎜⎝ 2
⎞ ⎟ ⎟ ⎠
1.4 / 0.4
Ma = 0.637
It yields
The speed of sound in air at the specified conditions is ⎛ 1000 m 2 / s 2 c = kRT = (1.4)(0.287 kJ/kg ⋅ K)(268.65 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 328.5 m/s ⎟ ⎠
Thus,
V = Ma × c = ( 0.637 )( 328.5 m/s) = 209 m/s Discussion
Note that the flow velocity can be measured in a simple and accurate way by simply measuring pressure.
12-140 Solution
The mass flow parameter m RT0 / ( AP0 ) versus the Mach number for k = 1.2, 1.4, and 1.6 in the range of
0 ≤ Ma ≤ 1 is to be plotted.
Analysis m RT0 P0 A
The mass flow rate parameter (m RT0 ) / P0 A can be expressed as ⎛ 2 = Ma k ⎜⎜ 2 ⎝ 2 + (k − 1) M
⎞ ⎟ ⎟ ⎠
( k +1) / 2 ( k −1)
0.75
Thus, Ma
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0
k = 1.2 0 0.1089 0.2143 0.3128 0.4015 0.4782 0.5411 0.5894 0.6230 0.6424 0.6485
k = 1.4 0 0.1176 0.2311 0.3365 0.4306 0.5111 0.5763 0.6257 0.6595 0.6787 0.6847
k = 1.6 1.4 1.2
0.60
k = 1.6 0 0.1257 0.2465 0.3582 0.4571 0.5407 0.6077 0.6578 0.6916 0.7106 0.7164
0.45
0.30
0.15
M 0
0.2
0.4
0.6
0.8
1.0
Discussion Note that the mass flow rate increases with increasing Mach number and specific heat ratio. It levels off at Ma = 1, and remains constant (choked flow).
12-79 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-141 Solution Helium gas is accelerated in a nozzle. The pressure and temperature of helium at the location where Ma = 1 and the ratio of the flow area at this location to the inlet flow area are to be determined. Assumptions 1 Helium is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic.
The properties of helium are R = 2.0769 kJ/kg⋅K, cp = 5.1926 kJ/kg⋅K, and k = 1.667.
Properties
Analysis The properties of the fluid at the location where Ma = 1 are the critical properties, denoted by superscript *. We first determine the stagnation temperature and pressure, which remain constant throughout the nozzle since the flow is isentropic. T0 = Ti +
Vi 2 (120 m/s) 2 ⎛ 1 kJ/kg ⎞ = 500 K + ⎜ ⎟ = 501.4 K 2c p 2 × 5.1926 kJ/kg ⋅ K ⎝ 1000 m 2 / s 2 ⎠ i
and
120 m/s
⎛T P0 = Pi ⎜⎜ 0 ⎝ Ti
⎞ ⎟ ⎟ ⎠
k /( k −1)
501.4 K ⎞ = (0.8 MPa)⎛⎜ ⎟ ⎝ 500 K ⎠
1.667 /(1.667 −1)
He
* Ma = 1
= 0.806 MPa
The Mach number at the nozzle exit is given to be Ma = 1. Therefore, the properties at the nozzle exit are the critical properties determined from 2 ⎛ 2 ⎞ ⎛ ⎞ T * = T0 ⎜ ⎟ = (501.4 K)⎜ ⎟ = 376 K ⎝ k +1⎠ ⎝ 1.667 + 1 ⎠ k /( k −1)
1.667 /(1.667 −1)
2 ⎛ 2 ⎞ ⎛ ⎞ P* = P0 ⎜ = (0.806 MPa)⎜ ⎟ ⎟ ⎝ k +1⎠ ⎝ 1.667 + 1 ⎠ The speed of sound and the Mach number at the nozzle inlet are
= 0.393 MPa
⎛ 1000 m 2 / s 2 c i = kRT i = (1.667)(2.0769 kJ/kg ⋅ K)(500 K)⎜⎜ ⎝ 1 kJ/kg Ma i =
⎞ ⎟ = 1316 m/s ⎟ ⎠
Vi 120 m/s = = 0.0912 c i 1316 m/s
The ratio of the entrance-to-throat area is 1 ⎡⎛ 2 ⎞⎛ k − 1 2 ⎞⎤ = ⎢⎜ k + 1 ⎟⎜1 + 2 Ma i ⎟⎥ * Ma ⎠⎝ ⎠⎦ A i ⎣⎝
Ai
( k +1) /[ 2 ( k −1)]
1 ⎡⎛ 2 ⎞⎛1 + 1.667 − 1 (0.0912) 2 ⎞⎤ ⎜ ⎟⎜ ⎟⎥ ⎢ 0.0912 ⎣⎝ 1.667 + 1 ⎠⎝ 2 ⎠⎦ = 6.20 Then the ratio of the throat area to the entrance area becomes A* 1 = = 0.161 Ai 6.20
2.667 /( 2×0.667 )
=
Discussion flow duct.
The compressible flow functions are essential tools when determining the proper shape of the compressible
12-80 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-142 Solution Helium gas enters a nozzle with negligible velocity, and is accelerated in a nozzle. The pressure and temperature of helium at the location where Ma = 1 and the ratio of the flow area at this location to the inlet flow area are to be determined. Assumptions 1 Helium is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. 3 The entrance velocity is negligible. Properties
The properties of helium are R = 2.0769 kJ/kg⋅K, cp = 5.1926 kJ/kg⋅K, and k = 1.667.
Analysis We treat helium as an ideal gas with k = 1.667. The properties of the fluid at the location where Ma = 1 are the critical properties, denoted by superscript *.
The stagnation temperature and pressure in this case are identical to the inlet temperature and pressure since the inlet velocity is negligible. They remain constant throughout the nozzle since the flow is isentropic. T0 = Ti = 500 K P0 = Pi = 0.8 MPa The Mach number at the nozzle exit is given to be Ma = 1. Therefore, the properties at the nozzle exit are the critical properties determined from 2 ⎛ 2 ⎞ ⎛ ⎞ T * = T0 ⎜ ⎟ = (500 K)⎜ ⎟ = 375 K ⎝ k +1 ⎠ ⎝ 1.667 + 1 ⎠ ⎛ 2 ⎞ P* = P0 ⎜ ⎟ ⎝ k +1⎠
k /( k −1)
2 ⎛ ⎞ = (0.8 MPa)⎜ ⎟ ⎝ 1.667 + 1 ⎠
1.667 /(1.667 −1)
= 0.390 MPa
i Vi ≈ 0
He
* Ma = 1
The ratio of the nozzle inlet area to the throat area is determined from 1 ⎡⎛ 2 ⎞⎛ k − 1 ⎞⎤ Ma i2 ⎟⎥ = ⎜ ⎟⎜1 + ⎢ * Ma i ⎣⎝ k + 1 ⎠⎝ 2 ⎠⎦ A
Ai
( k +1) /[ 2 ( k −1)]
But the Mach number at the nozzle inlet is Ma = 0 since Vi ≅ 0. Thus the ratio of the throat area to the nozzle inlet area is A* 1 = =0 ∞ Ai
Discussion flow duct.
The compressible flow functions are essential tools when determining the proper shape of the compressible
12-81 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-143
Solution Air enters a converging nozzle. The mass flow rate, the exit velocity, the exit Mach number, and the exit pressure-stagnation pressure ratio versus the back pressure-stagnation pressure ratio for a specified back pressure range are to be calculated and plotted. Assumptions 1 Air is an ideal gas with constant specific heats at room temperature. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. Properties The properties of air at room temperature are R = 0.287 kJ/kg⋅K, cp = 1.005 kJ/kg⋅K, and k = 1.4. Analysis We use EES to tabulate and plot the results. The stagnation properties remain constant throughout the nozzle since the flow is isentropic. They are determined from T0 = Ti +
Vi 2 (180 m/s) 2 ⎛ 1 kJ/kg ⎞ = 400 K + ⎜ ⎟ = 416.1 K 2c p 2 × 1.005 kJ/kg ⋅ K ⎝ 1000 m 2 / s 2 ⎠ k /( k −1)
1.4 /(1.4 −1) ⎛T ⎞ 416.1 K ⎞ = (900 kPa)⎛⎜ = 1033.3 kPa P0 = Pi ⎜⎜ 0 ⎟⎟ ⎟ ⎝ 400 K ⎠ ⎝ Ti ⎠ The critical pressure is determined to be k /( k −1)
i
Temperature
⎛P Te = T0 ⎜⎜ e ⎝ P0
1.4 / 0.4
⎞ ⎟ ⎟ ⎠
⎛ Pe ⎞ = (416.1 K)⎜ ⎟ ⎝ 1033.3 ⎠
⎛ 1000 m 2 / s 2 = (1.4)(0.287 kJ/kg ⋅ K)⎜⎜ ⎝ 1 kJ/kg
c e = kRT
Mach number
Ma e = Ve / c e
Density
ρe =
Mass flow rate
m = ρ eV e Ae = ρ eVe (0.001 m 2 )
e
Pe
Pb Ve
0.4 / 1.4
c
⎛ 1000 m 2 /s 2 Velocity V = 2c p (T0 − Te ) = 2(1.005 kJ/kg ⋅ K)(416.1 − Te )⎜⎜ ⎝ 1 kJ/kg Speed of sound
e
180 m/s
2 ⎞ 2 ⎞ P* = P0 ⎛⎜ = (1033.3 kPa)⎛⎜ = 545.9 kPa ⎟ ⎟ ⎝ k + 1⎠ ⎝ 1.4 + 1 ⎠ Then the pressure at the exit plane (throat) is for Pb ≥ 545.9 kPa Pe = Pb Pe = P* = 545.9 kPa for Pb < 545.9 kPa (choked flow) Thus the back pressure does not affect the flow when 100 < Pb < 545.9 kPa. For a specified exit pressure Pe, the temperature, velocity, and mass flow rate are ( k −1) / k
Air
⎞ ⎟ ⎟ ⎠
⎞ ⎟ ⎟ ⎠
Pb
m max m
Pe Pe = RTe (0.287 kPa ⋅ m 3 / kg ⋅ K )Te 100
475.5
900
Pb kPa
kg / s Ve, m/s Ma Pb, kPa Pb, P0 Pe, kPa Pb, P0 Te, K ρe, kg/m3 m, 900 0.871 900 0.871 400.0 180.0 0.45 7.840 0 800 0.774 800 0.774 386.8 162.9 0.41 7.206 1.174 700 0.677 700 0.677 372.3 236.0 0.61 6.551 1.546 600 0.581 600 0.581 356.2 296.7 0.78 5.869 1.741 545.9 0.528 545.9 0.528 333.3 366.2 1.00 4.971 1.820 500 0.484 545.9 0.528 333.2 366.2 1.00 4.971 1.820 400 0.387 545.9 0.528 333.3 366.2 1.00 4.971 1.820 300 0.290 545.9 0.528 333.3 366.2 1.00 4.971 1.820 200 0.194 545.9 0.528 333.3 366.2 1.00 4.971 1.820 100 0.097 545.9 0.528 333.3 366.2 1.00 4.971 1.820 Discussion Once the back pressure drops below 545.0 kPa, the flow is choked, and m remains constant from then on.
12-82 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-144
Solution Steam enters a converging nozzle. The exit pressure, the exit velocity, and the mass flow rate versus the back pressure for a specified back pressure range are to be plotted. Assumptions 1 Steam is to be treated as an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. 3 The nozzle is adiabatic.
The ideal gas properties of steam are R = 0.462 kJ/kg.K, cp = 1.872 kJ/kg.K, and k = 1.3.
Properties
Analysis We use EES to solve the problem. The stagnation properties in this case are identical to the inlet properties since the inlet velocity is negligible. Since the flow is isentropic, they remain constant throughout the nozzle, P0 = Pi = 6 MPa and T0 = Ti = 700 K The critical pressure is determined to be k /( k −1)
1.3 / 0.3
⎛ 2 ⎞ ⎛ 2 ⎞ P* = P0 ⎜ = (6 MPa)⎜ = 3.274 MPa ⎟ ⎟ ⎝ k + 1⎠ ⎝ 1.3 + 1 ⎠ Then the pressure at the exit plane (throat) is Pe = Pb for Pb ≥ 3.274 MPa Pe = P* = 3.274 MPa for Pb < 3.274 MPa (choked flow)
i
e
STEAM
Vi ≈ 0
Pe
Thus the back pressure does not affect the flow when 3 < Pb < 3.274 MPa. For a specified exit pressure Pe, the temperature, velocity, and mass flow rate are ⎛P Te = T0 ⎜⎜ e ⎝ P0
Temperature
⎞ ⎟ ⎟ ⎠
( k −1) / k
⎛P ⎞ = (700 K)⎜ e ⎟ ⎝ 6 ⎠
0.3 / 1.3
Pb
⎛ 1000 m 2 /s 2 Velocity V = 2c p (T0 − Te ) = 2(1.872 kJ/kg ⋅ K)(700 − Te )⎜⎜ ⎝ 1 kJ/kg P Pe ρe = e = Density RTe (0.462 kPa ⋅ m 3 / kg ⋅ K )Te Mass flow rate
⎞ ⎟ ⎟ ⎠
Discussion
C
Pb
m = ρ eV e Ae = ρ eVe (0.0008 m 2 ) m
The results of the calculations are tabulated as follows: Pb, MPa 6.0 5.5 5.0 4.5 4.0 3.5 3.274 3.0
Ve
Pe, MPa 6.0 5.5 5.0 4.5 4.0 3.5 3.274 3.274
Te, K 700 686.1 671.2 655.0 637.5 618.1 608.7 608.7
Ve, m/s 0 228.1 328.4 410.5 483.7 553.7 584.7 584.7
ρe, kg/m3 18.55 17.35 16.12 14.87 13.58 12.26 11.64 11.64
kg / s m, 0 3.166 4.235 4.883 5.255 5.431 5.445 5.445
max m
3
3.274
6
Pb MPa
Once the back pressure drops below 3.274 MPa, the flow is choked, and m remains constant from then on.
12-83 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-145 Solution An expression for the ratio of the stagnation pressure after a shock wave to the static pressure before the shock wave as a function of k and the Mach number upstream of the shock wave is to be found.
The relation between P1 and P2 is
Analysis
⎛ 1 + kMa 12 P2 1 + kMa 22 ⎜ P P = ⎯ ⎯→ = 2 1 ⎜ 1 + kMa 2 P1 1 + kMa 12 2 ⎝
⎞ ⎟ ⎟ ⎠
We substitute this into the isentropic relation k / ( k −1) P02 = (1 + ( k − 1) Ma 22 / 2 ) P2
which yields P02 ⎛ 1 + kMa 12 =⎜ P1 ⎜⎝ 1 + kMa 22
⎞ ⎟ 1 + (k − 1)Ma 22 / 2 ⎟ ⎠
(
)k /(k −1)
where Ma 22 =
Ma 12 + 2 /( k − 1) 2kMa 22 /( k − 1) − 1
Substituting, P02 ⎛ (1 + kMa 12 )(2kMa 12 − k + 1) ⎞⎛ (k − 1)Ma 12 / 2 + 1 ⎞ ⎟ ⎟⎜1 + = ⎜⎜ ⎟⎜ 2kMa 2 /( k − 1) − 1 ⎟ P1 ⎝ kMa 12 (k + 1) − k + 3 1 ⎠ ⎠⎝
Discussion
k /( k −1)
Similar manipulations of the equations can be performed to get the ratio of other parameters across a shock.
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Chapter 12 Compressible Flow 12-146 Solution Nitrogen entering a converging-diverging nozzle experiences a normal shock. The pressure, temperature, velocity, Mach number, and stagnation pressure downstream of the shock are to be determined. The results are to be compared to those of air under the same conditions. Assumptions 1 Nitrogen is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, onedimensional, and isentropic. 3 The nozzle is adiabatic. Properties
The properties of nitrogen are R = 0.297 kJ/kg⋅K and k = 1.4.
Analysis The inlet stagnation properties in this case are identical to the inlet properties since the inlet velocity is negligible. Assuming the flow before the shock to be isentropic, P01 = Pi = 700 kPa T01 = Ti = 300 K
shock wave
Then, ⎛ 2 T1 = T01 ⎜⎜ 2 ( 1)Ma 12 k + − ⎝
⎞ ⎛ 2 ⎟ = (300 K)⎜ ⎜ 2 + (1.4 - 1)3 2 ⎟ ⎝ ⎠
i
⎞ ⎟ = 107.1 K ⎟ ⎠
N2
1
2
Vi ≈ 0 Ma1 = 3.0
and ⎛T ⎞ P1 = P01 ⎜⎜ 1 ⎟⎟ ⎝ T01 ⎠
k /( k −1)
⎛ 107.1 ⎞ = (700 kPa)⎜ ⎟ ⎝ 300 ⎠
1.4 / 0.4
= 19.06 kPa
The fluid properties after the shock (denoted by subscript 2) are related to those before the shock through the functions listed in Table A-14. For Ma1 = 3.0 we read Ma 2 = 0.4752 ≅ 0.475 ,
P02 = 0.32834 , P01
P2 = 10.333, P1
and
T2 = 2.679 T1
Then the stagnation pressure P02, static pressure P2, and static temperature T2, are determined to be P02 = 0.32834 P01 = (0.32834)(700 kPa) = 230 kPa P2 = 10.333P1 = (10.333)(19.06 kPa) = 197 kPa T2 = 2.679T1 = (2.679)(107.1 K) = 287 K
The velocity after the shock can be determined from V2 = Ma2c2, where c2 is the speed of sound at the exit conditions after the shock, ⎛ 1000 m 2 / s 2 V 2 = Ma 2 c 2 = Ma 2 kRT2 = (0.4752) (1.4)(0.297 kJ/kg ⋅ K)(287 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 164 m/s ⎟ ⎠
Discussion For air at specified conditions k = 1.4 (same as nitrogen) and R = 0.287 kJ/kg·K. Thus the only quantity which will be different in the case of air is the velocity after the normal shock, which happens to be 161.3 m/s.
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Chapter 12 Compressible Flow 12-147 Solution The diffuser of an aircraft is considered. The static pressure rise across the diffuser and the exit area are to be determined. Assumptions 1 Air is an ideal gas with constant specific heats at room temperature. 2 Flow through the diffuser is steady, one-dimensional, and isentropic. 3 The diffuser is adiabatic. Properties
Air properties at room temperature are R = 0.287 kJ/kg⋅K, cp = 1.005 kJ/kg·K, and k = 1.4.
Analysis
The inlet velocity is
⎛ 1000 m 2 / s 2 V1 = Ma 1c1 = M 1 kRT1 = (0.8) (1.4)(0.287 kJ/kg ⋅ K)(242.7 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 249.8 m/s ⎟ ⎠
Then the stagnation temperature and pressure at the diffuser inlet become T01 = T1 +
V1 2 (249.8 m/s) 2 ⎛ 1 kJ/kg ⎞ = 242.7 + ⎜ ⎟ = 273.7 K 2c p 2(1.005 kJ/kg ⋅ K) ⎝ 1000 m 2 / s 2 ⎠
⎛T ⎞ P01 = P1 ⎜⎜ 01 ⎟⎟ ⎝ T1 ⎠
k /( k −1)
273.7 K ⎞ = (41.1 kPa)⎛⎜ ⎟ ⎝ 242.7 K ⎠
1.4 /(1.4 −1)
AIR 1 Ma1 = 0.8
2 Diffuser
Ma2 = 0.3
= 62.6 kPa
For an adiabatic diffuser, the energy equation reduces to h01 = h02. Noting that h = cpT and the specific heats are assumed to be constant, we have T01 = T02 = T0 = 273.7 K The isentropic relation between states 1 and 02 gives k /( k −1)
1.4 /(1.4 −1) ⎛T ⎞ ⎛ 273.72 K ⎞ = (41.1 kPa)⎜ = 62.61 kPa P02 = P02 = P1 ⎜⎜ 02 ⎟⎟ ⎟ ⎝ 242.7 K ⎠ ⎝ T1 ⎠ The exit velocity can be expressed as
⎛ 1000 m 2 / s 2 V 2 = Ma 2 c 2 = Ma 2 kRT2 = (0.3) (1.4)(0.287 kJ/kg ⋅ K) T2 ⎜⎜ ⎝ 1 kJ/kg
Thus
T2 = T02 −
6.012 T2 m 2 /s 2 ⎛ 1 kJ/kg V2 2 = (273.7) − ⎜ 2c p 2(1.005 kJ/kg ⋅ K) ⎝ 1000 m 2 / s 2
⎞ ⎟ = 6.01 T2 ⎟ ⎠
⎞ ⎟ = 268.9 K ⎠
Then the static exit pressure becomes k /( k −1)
1.4 /(1.4 −1) ⎛T ⎞ 268.9 K ⎞ = (62.61 kPa)⎛⎜ = 58.85 kPa P2 = P02 ⎜⎜ 2 ⎟⎟ ⎟ ⎝ 273.7 K ⎠ ⎝ T02 ⎠ Thus the static pressure rise across the diffuser is
Also,
ΔP = P2 − P1 = 58.85 − 41.1 = 17.8 kPa P 58.85 kPa ρ2 = 2 = = 0.7626 kg/m 3 RT2 (0.287 kPa ⋅ m 3 /kg ⋅ K)(268.9 K)
V 2 = 6.01 T2 = 6.01 268.9 = 98.6 m/s
Thus
A2 =
65 kg/s m = = 0.864 m 2 ρ 2V 2 (0.7626 kg/m 3 )(98.6 m/s)
The pressure rise in actual diffusers will be lower because of the irreversibilities. However, flow through Discussion well-designed diffusers is very nearly isentropic.
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Chapter 12 Compressible Flow 12-148 Solution Helium gas is accelerated in a nozzle isentropically. For a specified mass flow rate, the throat and exit areas of the nozzle are to be determined. Assumptions 1 Helium is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. 3 The nozzle is adiabatic. Properties The properties of helium are R = 2.0769 kJ/kg.K, cp = 5.1926 kJ/kg.K, and k = 1.667. Analysis The inlet stagnation properties in this case are identical to the inlet properties since the inlet velocity is negligible, T01 = T1 = 500 K P01 = P1 = 1.0 MPa
The flow is assumed to be isentropic, thus the stagnation temperature and pressure remain constant throughout the nozzle, T02 = T01 = 500 K
1
He
*
2
Vi ≈ 0
P02 = P01 = 1.0 MPa
The critical pressure and temperature are determined from 2 ⎞ 2 ⎛ ⎞ = 375.0 K T * = T0 ⎛⎜ ⎟ = (500 K)⎜ ⎟ ⎝ k + 1⎠ ⎝ 1.667 + 1 ⎠ k /( k −1)
1.667 /(1.667 −1)
2 ⎞ 2 ⎞ P* = P0 ⎛⎜ = (1.0 MPa)⎛⎜ = 0.487 MPa ⎟ ⎟ ⎝ k + 1⎠ ⎝ 1.667 + 1 ⎠ P* 487 kPa ρ* = = = 0.625 kg/m 3 RT * (2.0769 kPa ⋅ m 3 /kg ⋅ K)(375 K) ⎛ 1000 m 2 / s 2 ⎞ ⎟ = 1139.4 m/s V * = c* = kRT * = (1.667)(2.0769 kJ/kg ⋅ K)(375 K)⎜⎜ ⎟ ⎝ 1 kJ/kg ⎠ Thus the throat area is 0.25 kg/s m A* = = = 3.51× 10 − 4 m 2 = 3.51 cm 2 ρ *V * (0.625 kg/m 3 )(1139.4 m/s)
At the nozzle exit the pressure is P2 = 0.1 MPa. Then the other properties at the nozzle exit are determined to be k /( k −1) 1.667 / 0.667 P0 ⎛ k − 1 1.0 MPa ⎛ 1.667 − 1 ⎞ ⎞ = ⎜1 + Ma 22 ⎟ ⎯ ⎯→ = ⎜1 + Ma 22 ⎟ 2 0.1 MPa ⎝ 2 P2 ⎝ ⎠ ⎠
It yields Ma2 = 2.130, which is greater than 1. Therefore, the nozzle must be converging-diverging. ⎛ ⎞ ⎛ ⎞ 2 2 ⎟ = (500 K )⎜ ⎟ = 199.0 K T2 = T0 ⎜⎜ 2 ⎟ 2 ⎟ ⎜ ⎝ 2 + (1.667 − 1) × 2.13 ⎠ ⎝ 2 + (k − 1)Ma 2 ⎠ P 100 kPa ρ2 = 2 = = 0.242 kg/m 3 RT2 (2.0769 kPa ⋅ m 3 /kg ⋅ K)(199 K) ⎛ 1000 m 2 / s 2 V 2 = Ma 2 c 2 = Ma 2 kRT2 = (2.13) (1.667)(2.0769 kJ/kg ⋅ K)(199 K)⎜⎜ ⎝ 1 kJ/kg Thus the exit area is 0.25 kg/s m A2 = = = 5.84 × 10 − 4 m 2 = 5.84 cm 2 ρ 2V 2 (0.242 kg/m 3 )(1768 m/s) Discussion
⎞ ⎟ = 1768.0 m/s ⎟ ⎠
Flow areas in actual nozzles would be somewhat larger to accommodate the irreversibilities.
12-87 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-149E Solution Helium gas is accelerated in a nozzle. For a specified mass flow rate, the throat and exit areas of the nozzle are to be determined for the cases of isentropic and 97% efficient nozzles.
Assumptions 1 Helium is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. 3 The nozzle is adiabatic. Properties The properties of helium are R = 0.4961 Btu/lbm·R = 2.6809 psia·ft3/lbm·R, cp = 1.25 Btu/lbm·R, and k = 1.667. Analysis The inlet stagnation properties in this case are identical to the inlet properties since the inlet velocity is negligible, T01 = T1 = 900 R
P01 = P1 = 150 psia The flow is assumed to be isentropic, thus the stagnation temperature and pressure remain constant throughout the nozzle, T02 = T01 = 900 R P02 = P01 = 150 psia
1
He
*
2
Vi ≈ 0
The critical pressure and temperature are determined from 2 ⎞ 2 ⎛ ⎞ = 674.9 R T * = T0 ⎛⎜ ⎟ = (900 R)⎜ ⎟ ⎝ k + 1⎠ ⎝ 1.667 + 1 ⎠ k /( k −1)
1.667 /(1.667 −1)
2 ⎞ 2 ⎞ P* = P0 ⎛⎜ = (150 psia)⎛⎜ = 73.1 psia ⎟ ⎟ ⎝ k + 1⎠ ⎝ 1.667 + 1 ⎠ 73.1 psia P* ρ* = = = 0.0404 1bm/ft 3 RT * (2.6809 psia ⋅ ft 3 /lbm ⋅ R)(674.9 R)
and
⎛ 25,037 ft 2 / s 2 V * = c* = kRT * = (1.667)(0.4961 Btu/lbm ⋅ R)(674.9 R)⎜⎜ ⎝ 1 Btu/1bm m 0.2 1bm/s A* = = = 0.00132 ft 2 ρ *V * (0.0404 1bm/ft 3 )(3738 ft/s)
⎞ ⎟ = 3738 ft/s ⎟ ⎠
At the nozzle exit the pressure is P2 = 15 psia. Then the other properties at the nozzle exit are determined to be k /( k −1) 1.667 / 0.667 p0 ⎛ k − 1 150 psia ⎛ 1.667 − 1 ⎞ ⎞ = ⎜1 + Ma 22 ⎟ ⎯ ⎯→ = ⎜1 + Ma 22 ⎟ 2 15 psia ⎝ 2 p2 ⎝ ⎠ ⎠ It yields Ma2 = 2.130, which is greater than 1. Therefore, the nozzle must be converging-diverging. ⎛ ⎞ ⎛ ⎞ 2 2 ⎟ = (900 R )⎜ ⎟ = 358.1 R T2 = T0 ⎜⎜ 2 ⎟ 2 ⎟ ⎜ ⎝ 2 + (1.667 − 1) × 2.13 ⎠ ⎝ 2 + (k − 1)Ma 2 ⎠ P 15 psia ρ2 = 2 = = 0.0156 1bm/ft 3 3 RT2 (2.6809 psia ⋅ ft /lbm ⋅ R)(358.1 R) ⎛ 25,037 ft 2 / s 2 V 2 = Ma 2 c 2 = Ma 2 kRT2 = (2.13) (1.667)(0.4961 Btu/lbm ⋅ R)(358.1 R)⎜⎜ ⎝ 1 Btu/1bm Thus the exit area is m 0.2 lbm/s A2 = = = 0.00221 ft 2 ρ 2V 2 (0.0156 lbm/ft 3 )(5800 ft/s)
Discussion
⎞ ⎟ = 5800 ft/s ⎟ ⎠
Flow areas in actual nozzles would be somewhat larger to accommodate the irreversibilities.
12-88 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-150 [Also solved using EES on enclosed DVD]
Solution Using the compressible flow relations, the one-dimensional compressible flow functions are to be evaluated and tabulated as in Table A-13 for an ideal gas with k = 1.667. Properties Analysis Ma * = Ma
The specific heat ratio of the ideal gas is given to be k = 1.667. The compressible flow functions listed below are expressed in EES and the results are tabulated. k +1
1 ⎡⎛ 2 ⎞⎛ k − 1 2 ⎞⎤ = ⎢⎜ k + 1 ⎟⎜1 + 2 Ma ⎟⎥ * Ma ⎠⎝ ⎠⎦ A ⎣⎝
A
2 + (k − 1)Ma 2
P ⎛ k −1 ⎞ Ma 2 ⎟ = ⎜1 + P0 ⎝ 2 ⎠ T ⎛ k −1 ⎞ Ma 2 ⎟ = ⎜1 + 2 T0 ⎝ ⎠
− k /( k −1)
0.5( k +1) /( k −1)
−1 /( k −1) ρ ⎛ k −1 ⎞ Ma 2 ⎟ = ⎜1 + 2 ρ0 ⎝ ⎠
−1
k=1.667 PP0=(1+(k-1)*M^2/2)^(-k/(k-1)) TT0=1/(1+(k-1)*M^2/2) DD0=(1+(k-1)*M^2/2)^(-1/(k-1)) Mcr=M*SQRT((k+1)/(2+(k-1)*M^2)) AAcr=((2/(k+1))*(1+0.5*(k-1)*M^2))^(0.5*(k+1)/(k-1))/M
Ma
Ma*
A/A*
P/P0
T/T0
0 0.1153 0.2294 0.3413 0.4501 0.5547 0.6547 0.7494 0.8386 0.9222 1.0000 1.1390 1.2572 1.3570 1.4411 1.5117 1.5713 1.6216 1.6643 1.7007 1.7318 1.8895 1.9996
∞ 5.6624 2.8879 1.9891 1.5602 1.3203 1.1760 1.0875 1.0351 1.0081 1.0000 1.0267 1.0983 1.2075 1.3519 1.5311 1.7459 1.9980 2.2893 2.6222 2.9990 9.7920 ∞
ρ/ρ0
0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0 5.0 ∝
1.0000 0.9917 0.9674 0.9288 0.8782 0.8186 0.7532 0.6850 0.6166 0.5501 0.4871 0.3752 0.2845 0.2138 0.1603 0.1202 0.0906 0.0686 0.0524 0.0403 0.0313 0.0038 0
1.0000 0.9950 0.9803 0.9566 0.9250 0.8869 0.8437 0.7970 0.7482 0.6987 0.6495 0.5554 0.4704 0.3964 0.3334 0.2806 0.2368 0.2005 0.1705 0.1457 0.1251 0.0351 0
1.0000 0.9967 0.9868 0.9709 0.9493 0.9230 0.8928 0.8595 0.8241 0.7873 0.7499 0.6756 0.6047 0.5394 0.4806 0.4284 0.3825 0.3424 0.3073 0.2767 0.2499 0.1071 0
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.667.
12-89 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-151 [Also solved using EES on enclosed DVD]
Solution Using the normal shock relations, the normal shock functions are to be evaluated and tabulated as in Table A-14 for an ideal gas with k = 1.667. Properties Analysis
The specific heat ratio of the ideal gas is given to be k = 1.667. The normal shock relations listed below are expressed in EES and the results are tabulated.
Ma 2 =
P2 1 + kMa 12 2kMa 12 − k + 1 = = P1 1 + kMa 22 k +1
(k − 1)Ma 12 + 2 2kMa 12 − k + 1
(k + 1)Ma 12 ρ 2 P2 / P1 V = = = 1 , ρ 1 T2 / T1 2 + (k − 1)Ma 12 V 2
T2 2 + Ma 12 (k − 1) = T1 2 + Ma 22 (k − 1) k +1
P02 Ma 1 ⎡1 + Ma 22 (k − 1) / 2 ⎤ 2( k −1) = ⎢ ⎥ P01 Ma 2 ⎢⎣1 + Ma 12 (k − 1) / 2 ⎦⎥
P02 (1 + kMa 12 )[1 + Ma 22 (k − 1) / 2] k /( k −1) = P1 1 + kMa 22
k=1.667 My=SQRT((Mx^2+2/(k-1))/(2*Mx^2*k/(k-1)-1)) PyPx=(1+k*Mx^2)/(1+k*My^2) TyTx=(1+Mx^2*(k-1)/2)/(1+My^2*(k-1)/2) RyRx=PyPx/TyTx P0yP0x=(Mx/My)*((1+My^2*(k-1)/2)/(1+Mx^2*(k-1)/2))^(0.5*(k+1)/(k-1)) P0yPx=(1+k*Mx^2)*(1+My^2*(k-1)/2)^(k/(k-1))/(1+k*My^2)
Ma1 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 4.0 5.0 ∞
Ma2 1.0000 0.9131 0.8462 0.7934 0.7508 0.7157 0.6864 0.6618 0.6407 0.6227 0.6070 0.5933 0.5814 0.5708 0.5614 0.5530 0.5455 0.5388 0.5327 0.5273 0.5223 0.4905 0.4753 0.4473
P2/P1 1.0000 1.2625 1.5500 1.8626 2.2001 2.5626 2.9501 3.3627 3.8002 4.2627 4.7503 5.2628 5.8004 6.3629 6.9504 7.5630 8.2005 8.8631 9.5506 10.2632 11.0007 19.7514 31.0022 ∞
ρ2/ρ1 1.0000 1.1496 1.2972 1.4413 1.5805 1.7141 1.8415 1.9624 2.0766 2.1842 2.2853 2.3802 2.4689 2.5520 2.6296 2.7021 2.7699 2.8332 2.8923 2.9476 2.9993 3.3674 3.5703 3.9985
T2/T1 1.0000 1.0982 1.1949 1.2923 1.3920 1.4950 1.6020 1.7135 1.8300 1.9516 2.0786 2.2111 2.3493 2.4933 2.6432 2.7989 2.9606 3.1283 3.3021 3.4819 3.6678 5.8654 8.6834 ∞
P02/P01 1 0.999 0.9933 0.9813 0.9626 0.938 0.9085 0.8752 0.8392 0.8016 0.763 0.7243 0.6861 0.6486 0.6124 0.5775 0.5442 0.5125 0.4824 0.4541 0.4274 0.2374 0.1398 0
P02/P1 2.0530 2.3308 2.6473 2.9990 3.3838 3.8007 4.2488 4.7278 5.2371 5.7767 6.3462 6.9457 7.5749 8.2339 8.9225 9.6407 10.3885 11.1659 11.9728 12.8091 13.6750 23.9530 37.1723 ∞
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.667.
12-90 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-152 Solution The critical temperature, pressure, and density of an equimolar mixture of oxygen and nitrogen for specified stagnation properties are to be determined.
Assumptions
Both oxygen and nitrogen are ideal gases with constant specific heats at room temperature.
Properties The specific heat ratio and molar mass are k = 1.395 and M = 32 kg/kmol for oxygen, and k = 1.4 and M = 28 kg/kmol for nitrogen. Analysis
The gas constant of the mixture is
M m = y O2 M O2 + y N 2 M N 2 = 0.5 × 32 + 0.5 × 28 = 30 kg/kmol Rm =
8.314 kJ/kmol ⋅ K Ru = = 0.2771 kJ/kg ⋅ K 30 kg/kmol Mm
The specific heat ratio is 1.4 for nitrogen, and nearly 1.4 for oxygen. Therefore, the specific heat ratio of the mixture is also 1.4. Then the critical temperature, pressure, and density of the mixture become ⎛ 2 ⎞ ⎛ 2 ⎞ T * = T0 ⎜ ⎟ = (800 K)⎜ ⎟ = 667 K ⎝ k +1⎠ ⎝ 1.4 + 1 ⎠ k /( k −1)
1.4 /(1.4 −1)
⎛ 2 ⎞ ⎛ 2 ⎞ = (500 kPa)⎜ = 264 kPa P* = P0 ⎜ ⎟ ⎟ ⎝ k +1⎠ ⎝ 1.4 + 1 ⎠ 264 kPa P* ρ* = = = 1.43 kg/m 3 RT * (0.2771 kPa ⋅ m 3 /kg ⋅ K)(667 K)
Discussion If the specific heat ratios k of the two gases were different, then we would need to determine the k of the mixture from k = Cp,m/Cv,m where the specific heats of the mixture are determined from
C p ,m = mf O2 C p ,O2 + mf N 2 C p , N 2 = ( y O2 M O2 / M m )C p ,O2 + ( y N 2 M N 2 / M m )C p , N 2 C v ,m = mf O2 C v ,O2 + mf N 2 C v , N 2 = ( y O2 M O2 / M m )C v ,O2 + ( y N 2 M N 2 / M m )C v, N 2
where mf is the mass fraction and y is the mole fraction. In this case it would give C p ,m = (0.5 × 32 / 30) × 0.918 + (0.5 × 28 / 30) × 1.039 = 0.974 kJ/kg.K C p ,m = (0.5 × 32 / 30) × 0.658 + (0.5 × 28 / 30) × 0.743 = 0.698 kJ/kg.K
and k = 0.974/0.698 = 1.40
12-91 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-153
Solution Using EES (or other) software, the shape of a converging-diverging nozzle is to be determined for specified flow rate and stagnation conditions. The nozzle and the Mach number are to be plotted. Assumptions 1 Air is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady, one-dimensional, and isentropic. 3 The nozzle is adiabatic. Properties
The specific heat ratio of air at room temperature is 1.4.
Analysis
The problem is solved using EES, and the results are tabulated and plotted below.
k=1.4 Cp=1.005 "kJ/kg.K" R=0.287 "kJ/kg.K" P0=1400 "kPa" T0=200+273 "K" m=3 "kg/s" rho_0=P0/(R*T0) rho=P/(R*T) T=T0*(P/P0)^((k-1)/k) V=SQRT(2*Cp*(T0-T)*1000) A=m/(rho*V)*10000 "cm2" C=SQRT(k*R*T*1000) Ma=V/C
2.5
2
Ma
1.5
1
0.5
Mach number Ma
1400 1350 1300 1250 1200 1150 1100 1050 1000 950 900 850 800 750 700 650 600 550 500 450 400 350 300 250 200 150 100
∞ 30.1 21.7 18.1 16.0 14.7 13.7 13.0 12.5 12.2 11.9 11.7 11.6 11.5 11.5 11.6 11.8 12.0 12.3 12.8 13.3 14.0 15.0 16.4 18.3 21.4 27.0
0 0.229 0.327 0.406 0.475 0.538 0.597 0.655 0.710 0.766 0.820 0.876 0.931 0.988 1.047 1.107 1.171 1.237 1.308 1.384 1.467 1.559 1.663 1.784 1.929 2.114 2.373
0 0
200
400
600
800
1000
1200
1400
1000
1200
1400
P, kPa
50 45 40 35 2
Flow area A, cm2
Flow area A, cm
Pressure P, kPa
30 25 20 15 10 0
200
400
600
800
P, kPa
Discussion The shape is not actually to scale since the horizontal axis is pressure rather than distance. If the pressure decreases linearly with distance, then the shape would be to scale.
12-92 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-154
Solution Using the compressible flow relations, the one-dimensional compressible flow functions are to be evaluated and tabulated as in Table A-13 for air. Properties The specific heat ratio is given to be k = 1.4 for air. Analysis Ma * = Ma
The compressible flow functions listed below are expressed in EES and the results are tabulated. k +1
A
2 + (k − 1)Ma 2
A*
P ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ 2 P0 ⎝ ⎠ T ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ T0 ⎝ 2 ⎠
− k /( k −1)
=
1 Ma
⎡⎛ 2 ⎞⎛ k − 1 2 ⎞⎤ ⎢⎜ k + 1 ⎟⎜1 + 2 Ma ⎟⎥ ⎠⎝ ⎠⎦ ⎣⎝
0.5( k +1) /( k −1)
−1 /( k −1) ρ ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ 2 ρ0 ⎝ ⎠
−1
Air: k=1.4 PP0=(1+(k-1)*M^2/2)^(-k/(k-1)) TT0=1/(1+(k-1)*M^2/2) DD0=(1+(k-1)*M^2/2)^(-1/(k-1)) Mcr=M*SQRT((k+1)/(2+(k-1)*M^2)) AAcr=((2/(k+1))*(1+0.5*(k-1)*M^2))^(0.5*(k+1)/(k-1))/M
Ma 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0
Ma* 1.0000 1.3646 1.6330 1.8257 1.9640 2.0642 2.1381 2.1936 2.2361 2.2691 2.2953 2.3163 2.3333 2.3474 2.3591 2.3689 2.3772 2.3843 2.3905
A/A* 1.0000 1.1762 1.6875 2.6367 4.2346 6.7896 10.7188 16.5622 25.0000 36.8690 53.1798 75.1343 104.1429 141.8415 190.1094 251.0862 327.1893 421.1314 535.9375
P/P0 0.5283 0.2724 0.1278 0.0585 0.0272 0.0131 0.0066 0.0035 0.0019 0.0011 0.0006 0.0004 0.0002 0.0002 0.0001 0.0001 0.0000 0.0000 0.0000
ρ/ρ0 0.6339 0.3950 0.2300 0.1317 0.0762 0.0452 0.0277 0.0174 0.0113 0.0076 0.0052 0.0036 0.0026 0.0019 0.0014 0.0011 0.0008 0.0006 0.0005
T/T0 0.8333 0.6897 0.5556 0.4444 0.3571 0.2899 0.2381 0.1980 0.1667 0.1418 0.1220 0.1058 0.0926 0.0816 0.0725 0.0647 0.0581 0.0525 0.0476
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.4.
12-93 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-155
Solution Using the compressible flow relations, the one-dimensional compressible flow functions are to be evaluated and tabulated as in Table A-13 for methane. Properties The specific heat ratio is given to be k = 1.3 for methane. Analysis Ma * = Ma
The compressible flow functions listed below are expressed in EES and the results are tabulated. k +1
A
2 + (k − 1)Ma 2
A*
P ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ 2 P0 ⎝ ⎠ T ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ T0 ⎝ 2 ⎠
− k /( k −1)
=
1 Ma
⎡⎛ 2 ⎞⎛ k − 1 2 ⎞⎤ ⎢⎜ k + 1 ⎟⎜1 + 2 Ma ⎟⎥ ⎠⎝ ⎠⎦ ⎣⎝
0.5( k +1) /( k −1)
−1 /( k −1) ρ ⎛ k −1 ⎞ = ⎜1 + Ma 2 ⎟ 2 ρ0 ⎝ ⎠
−1
Methane: k=1.3 PP0=(1+(k-1)*M^2/2)^(-k/(k-1)) TT0=1/(1+(k-1)*M^2/2) DD0=(1+(k-1)*M^2/2)^(-1/(k-1)) Mcr=M*SQRT((k+1)/(2+(k-1)*M^2)) AAcr=((2/(k+1))*(1+0.5*(k-1)*M^2))^(0.5*(k+1)/(k-1))/M
Ma 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0
Ma* 1.0000 1.3909 1.6956 1.9261 2.0986 2.2282 2.3263 2.4016 2.4602 2.5064 2.5434 2.5733 2.5978 2.6181 2.6350 2.6493 2.6615 2.6719 2.6810
A/A* 1.0000 1.1895 1.7732 2.9545 5.1598 9.1098 15.9441 27.3870 45.9565 75.2197 120.0965 187.2173 285.3372 425.8095 623.1235 895.5077 1265.6040 1761.2133 2416.1184
P/P0 0.5457 0.2836 0.1305 0.0569 0.0247 0.0109 0.0050 0.0024 0.0012 0.0006 0.0003 0.0002 0.0001 0.0001 0.0000 0.0000 0.0000 0.0000 0.0000
ρ/ρ0 0.6276 0.3793 0.2087 0.1103 0.0580 0.0309 0.0169 0.0095 0.0056 0.0033 0.0021 0.0013 0.0008 0.0006 0.0004 0.0003 0.0002 0.0001 0.0001
T/T0 0.8696 0.7477 0.6250 0.5161 0.4255 0.3524 0.2941 0.2477 0.2105 0.1806 0.1563 0.1363 0.1198 0.1060 0.0943 0.0845 0.0760 0.0688 0.0625
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.3.
12-94 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-156
Solution A-14 for air. Properties
The specific heat ratio is given to be k = 1.4 for air.
Analysis
The normal shock relations listed below are expressed in EES and the results are tabulated.
Using the normal shock relations, the normal shock functions are to be evaluated and tabulated as in Table
Ma 2 =
P2 1 + kMa 12 2kMa 12 − k + 1 = = P1 1 + kMa 22 k +1
(k − 1)Ma 12 + 2 2kMa 12 − k + 1
(k + 1)Ma 12 V ρ 2 P2 / P1 = = = 1 , ρ 1 T2 / T1 2 + (k − 1)Ma 12 V 2
T2 2 + Ma 12 (k − 1) = T1 2 + Ma 22 (k − 1) k +1
P02 Ma 1 ⎡1 + Ma 22 (k − 1) / 2 ⎤ 2( k −1) = ⎢ ⎥ P01 Ma 2 ⎢⎣1 + Ma 12 (k − 1) / 2 ⎦⎥
P02 (1 + kMa 12 )[1 + Ma 22 (k − 1) / 2] k /( k −1) = P1 1 + kMa 22
Air: k=1.4 My=SQRT((Mx^2+2/(k-1))/(2*Mx^2*k/(k-1)-1)) PyPx=(1+k*Mx^2)/(1+k*My^2) TyTx=(1+Mx^2*(k-1)/2)/(1+My^2*(k-1)/2) RyRx=PyPx/TyTx P0yP0x=(Mx/My)*((1+My^2*(k-1)/2)/(1+Mx^2*(k-1)/2))^(0.5*(k+1)/(k-1)) P0yPx=(1+k*Mx^2)*(1+My^2*(k-1)/2)^(k/(k-1))/(1+k*My^2)
Ma1 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0
Ma2 1.0000 0.7011 0.5774 0.5130 0.4752 0.4512 0.4350 0.4236 0.4152 0.4090 0.4042 0.4004 0.3974 0.3949 0.3929 0.3912 0.3898 0.3886 0.3876
P2/P1 1.0000 2.4583 4.5000 7.1250 10.3333 14.1250 18.5000 23.4583 29.0000 35.1250 41.8333 49.1250 57.0000 65.4583 74.5000 84.1250 94.3333 105.1250 116.5000
ρ2/ρ1 1.0000 1.8621 2.6667 3.3333 3.8571 4.2609 4.5714 4.8119 5.0000 5.1489 5.2683 5.3651 5.4444 5.5102 5.5652 5.6117 5.6512 5.6850 5.7143
T2/T1 1.0000 1.3202 1.6875 2.1375 2.6790 3.3151 4.0469 4.8751 5.8000 6.8218 7.9406 9.1564 10.4694 11.8795 13.3867 14.9911 16.6927 18.4915 20.3875
P02/P01 1 0.9298 0.7209 0.499 0.3283 0.2129 0.1388 0.0917 0.06172 0.04236 0.02965 0.02115 0.01535 0.01133 0.008488 0.006449 0.004964 0.003866 0.003045
P02/P1 1.8929 3.4133 5.6404 8.5261 12.0610 16.2420 21.0681 26.5387 32.6535 39.4124 46.8152 54.8620 63.5526 72.8871 82.8655 93.4876 104.7536 116.6634 129.2170
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.4.
12-95 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-157
Solution Using the normal shock relations, the normal shock functions are to be evaluated and tabulated as in Table A-14 for methane. Properties The specific heat ratio is given to be k = 1.3 for methane.
The normal shock relations listed below are expressed in EES and the results are tabulated.
Analysis
Ma 2 =
P2 1 + kMa 12 2kMa 12 − k + 1 = = P1 1 + kMa 22 k +1
(k − 1)Ma 12 + 2 2kMa 12 − k + 1
(k + 1)Ma 12 V ρ 2 P2 / P1 = = = 1 , ρ 1 T2 / T1 2 + (k − 1)Ma 12 V 2
T2 2 + Ma 12 (k − 1) = T1 2 + Ma 22 (k − 1) k +1
P02 Ma 1 ⎡1 + Ma 22 (k − 1) / 2 ⎤ 2( k −1) = ⎢ ⎥ P01 Ma 2 ⎢⎣1 + Ma 12 (k − 1) / 2 ⎦⎥
P02 (1 + kMa 12 )[1 + Ma 22 (k − 1) / 2] k /( k −1) = P1 1 + kMa 22
Methane: k=1.3 My=SQRT((Mx^2+2/(k-1))/(2*Mx^2*k/(k-1)-1)) PyPx=(1+k*Mx^2)/(1+k*My^2) TyTx=(1+Mx^2*(k-1)/2)/(1+My^2*(k-1)/2) RyRx=PyPx/TyTx P0yP0x=(Mx/My)*((1+My^2*(k-1)/2)/(1+Mx^2*(k-1)/2))^(0.5*(k+1)/(k-1)) P0yPx=(1+k*Mx^2)*(1+My^2*(k-1)/2)^(k/(k-1))/(1+k*My^2)
Ma1 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0
Ma2 1.0000 0.6942 0.5629 0.4929 0.4511 0.4241 0.4058 0.3927 0.3832 0.3760 0.3704 0.3660 0.3625 0.3596 0.3573 0.3553 0.3536 0.3522 0.3510
P2/P1 1.0000 2.4130 4.3913 6.9348 10.0435 13.7174 17.9565 22.7609 28.1304 34.0652 40.5652 47.6304 55.2609 63.4565 72.2174 81.5435 91.4348 101.8913 112.9130
ρ2/ρ1 1.0000 1.9346 2.8750 3.7097 4.4043 4.9648 5.4118 5.7678 6.0526 6.2822 6.4688 6.6218 6.7485 6.8543 6.9434 7.0190 7.0837 7.1393 7.1875
T2/T1 1.0000 1.2473 1.5274 1.8694 2.2804 2.7630 3.3181 3.9462 4.6476 5.4225 6.2710 7.1930 8.1886 9.2579 10.4009 11.6175 12.9079 14.2719 15.7096
P02/P01 1 0.9261 0.7006 0.461 0.2822 0.1677 0.09933 0.05939 0.03613 0.02243 0.01422 0.009218 0.006098 0.004114 0.002827 0.001977 0.001404 0.001012 0.000740
P02/P1 1.8324 3.2654 5.3700 8.0983 11.4409 15.3948 19.9589 25.1325 30.9155 37.3076 44.3087 51.9188 60.1379 68.9658 78.4027 88.4485 99.1032 110.367 122.239
Discussion The tabulated values are useful for quick calculations, but be careful – they apply only to one specific value of k, in this case k = 1.3.
12-96 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-158 Solution Air enters a constant-area adiabatic duct at a specified state, and leaves at a specified pressure. The mass flow rate of air, the exit velocity, and the average friction factor are to be determined.
P1= 97 kPa
P2 = 55 kPa
P0 = 100 kPa T0 = 300 K
D =3 cm
T2 Vacuum V2 pump
L=2m Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor is constant along the duct. Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K. The friction factor is given to be f = 0.025. Analysis Noting that the flow in the nozzle section is isentropic, the Mach number, thermodynamic temperature, and density at the tube inlet become − k /( k −1) −1.4 / 0.4 ⎛ k −1 ⎞ ⎛ 1.4 - 1 ⎞ P1 = P01 ⎜1 + Ma 12 ⎟ → 97 kPa = (100 kPa)⎜1 + Ma 12 ⎟ → Ma1 = 0.2091 2 2 ⎝ ⎠ ⎝ ⎠ −1
⎛ k −1 ⎞ ⎛ 1.4 - 1 ⎞ Ma 12 ⎟ = (300 K)⎜1 + T1 = T01 ⎜1 + (0.2091) 2 ⎟ 2 2 ⎝ ⎠ ⎝ ⎠ P 97 kPa = 1.136 kg/m 3 ρ1 = 1 = RT1 (0.287 kJ/kgK)(297.4 K)
−1
= 297.4 K
Then the inlet velocity and the mass flow rate become ⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(297.4 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 345.7 m/s ⎟ ⎠
V1 = Ma 1c1 = 0.2091(345.7 m/s ) = 72.3 m/s m air = ρ1 Ac1V1 = (1.136 kg/m 3 )[π (0.03 m) 2 / 4](72.3 m/s) = 0.0581 kg/s
The Fanno flow functions corresponding to the inlet Mach number are, from Table A-16, Ma1 = 0.2091:
(fL*/Dh)1 =13.095
T1/T* = 1.1896, P1/P* = 5.2173, V1/V* = 0.2280
Therefore,P1 = 5.2173P*. Then the Fanno function P2/P* becomes P2
P2 5.2173(55 kPa ) = = 2.9583 P1 / 5.2173 97 kPa P The corresponding Mach number and Fanno flow functions are, from Table A-16, *
=
Ma2 = 0.3655,
(fL*/Dh)1 = 3.0420, and
V2/V* =0.3951.
Then the air velocity at the duct exit and the average friction factor become V 2 V 2 / V * 0.3951 = = = 1.7329 V 1 V1 / V * 0.2280
→ V 2 = 1.7329V1 = 1.7329 (72.3 m/s ) = 125 m/s
⎛ fL* fL* ⎞ D 0.03 m L = L*1 − L*2 = ⎜⎜ 1 − 2 ⎟⎟ h → 2 m = (13.095 − 3.042) → f = 0.151 D D f f h ⎠ ⎝ h
Discussion Note that the mass flow rate and the average friction factor can be determined by measuring static pressure, as in incompressible flow.
12-97 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-159 Solution Supersonic airflow in a constant cross-sectional area adiabatic duct is considered. For a specified exit Mach number, the temperature, pressure, and velocity at the duct exit are to be determined.
L P1 = 80 kPa T1 = 250 K Ma1 = 2.2
L2* f = 0.03 Exit Ma2= 1.8
Ma*=1 T* P* V*
L1* Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor remains constant along the duct. Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K. The average friction factor is given to be f = 0.03.
The inlet velocity is
Analysis
⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(250 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 316.9 m/s ⎟ ⎠
V1 = Ma 1 c1 = 2.2(316.9 m/s) = 697.3 m/s The Fanno flow functions corresponding to the inlet and exit Mach numbers are, from Table A-16,
Ma1 = 2.2: Ma2 = 1.8:
(fL*/Dh)1 = 0.3609 (fL*/Dh)2 = 0.2419
T1/T* = 0.6098, P1/P* = 0.3549, V1/V* = 1.7179 T2/T* = 0.7282, P2/P* = 0.4741, V2/V* = 1.5360
Then the temperature, pressure, and velocity at the duct exit are determined to be T2 T2 / T * 0.7282 = = = 1.1942 T1 T1 / T * 0.6098
→ T2 = 1.1942T1 = 1.1942(250 K ) = 299 K
P2 P2 / P * 0.4741 = = = 1.3359 P1 P1 / P * 0.3549
→ P2 = 1.3359 P1 = 1.3359(80 kPa ) = 107 kPa
V 2 V 2 / V * 1.5360 = = = 0.8941 V 1 V1 / V * 1.7179
→ V 2 = 0.8941V1 = 0.8941(697.3 m/s) = 623 m/s
Discussion
The duct length is determined to be
⎛ fL* fL* ⎞ D 0.04 m L = L*1 − L*2 = ⎜⎜ 1 − 2 ⎟⎟ h = (0.3609 − 0.2419) = 0.16 m D D f 0.03 h ⎠ ⎝ h
Note that it takes a duct length of only 0.16 m for the Mach number to decrease from 2.2 to 1.8. The maximum (or sonic) duct lengths at the inlet and exit states in this case are L1* = 0.48 m and L2* = 0.32 m. Therefore, the flow would reach sonic conditions if a 0.32-m long section were added to the existing duct.
12-98 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-160 Solution Air flowing at a supersonic velocity in a duct is accelerated by cooling. For a specified exit Mach number, the rate of heat transfer is to be determined.
Q
Assumptions The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid.
P01 = 240 kPa T01 = 350 K
Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K.
Ma2 = 2
Ma1 = 1.2
Analysis Knowing stagnation properties, the static properties are determined to be
⎛ k −1 ⎞ T1 = T01 ⎜1 + Ma 12 ⎟ 2 ⎝ ⎠
−1
⎛ k −1 ⎞ Ma 12 ⎟ P1 = P01 ⎜1 + 2 ⎝ ⎠
− k /( k −1)
ρ1 =
⎛ 1.4 - 1 2 ⎞ 1.2 ⎟ = (350 K)⎜1 + 2 ⎝ ⎠
−1
= 271.7 K
⎛ 1.4 - 1 2 ⎞ = (240 kPa)⎜1 + 1.2 ⎟ 2 ⎝ ⎠
−1.4 / 0.4
= 98.97 kPa
P1 98.97 kPa = = 1.269 kg/m 3 RT1 (0.287 kJ/kgK)(271.7 K)
Then the inlet velocity and the mass flow rate become ⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(271.7 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 330.4 m/s ⎟ ⎠
V1 = Ma 1c1 = 1.2(330.4 m/s ) = 396.5 m/s m air = ρ1 Ac1V1 = (1.269 kg/m 3 )[π (0.20 m) 2 / 4](396.5 m/s) = 15.81 kg/s
The Rayleigh flow functions T0/T0* corresponding to the inlet and exit Mach numbers are (Table A-15): Ma1 = 1.8: Ma2 = 2:
T01/T0* = 0.9787 T02/T0* = 0.7934
Then the exit stagnation temperature is determined to be T0 2 T02 / T0* 0.7934 = = = 0.8107 T0 1 T01 / T0* 0.9787
→ T02 = 0.8107T01 = 0.8107(350 K ) = 283.7 K
Finally, the rate of heat transfer is Q = m air c p (T02 − T01 ) = (15.81 kg/s )(1.005 kJ/kg ⋅ K )( 283.7 − 350 ) K = −1053 kW ≅ −1050 kW Discussion The negative sign confirms that the gas needs to be cooled in order to be accelerated. Also, it can be shown that the thermodynamic temperature drops to 158 K at the exit, which is extremely low. Therefore, the duct may need to be heavily insulated to maintain indicated flow conditions.
12-99 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-161 Solution Air flowing at a subsonic velocity in a duct is accelerated by heating. The highest rate of heat transfer without affecting the inlet conditions is to be determined.
Assumptions 1 The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. 2 Inlet conditions (and thus the mass flow rate) remain constant. Properties
We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K.
Analysis Heat transfer will stop when the flow is choked, and thus Ma2 = V2/c2 = 1. The inlet density and stagnation temperature are Q P1 400 kPa = = 3.871 kg/m 3 ρ1 = RT1 (0.287 kJ/kgK)(360 K) P1 = 400 kPa T1 = 360 K Ma2 = 1 ⎛ k −1 ⎞ ⎛ 1.4 - 1 ⎞ T01 = T1 ⎜1 + Ma 12 ⎟ = (360 K)⎜1 + 0.4 2 ⎟ = 371.5 K 2 2 ⎝ ⎠ ⎝ ⎠ Ma1 = 0.4
Then the inlet velocity and the mass flow rate become ⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4 )(0.287 kJ/kg ⋅ K)(360 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 380.3 m/s ⎟ ⎠
V1 = Ma 1c1 = 0.4(380.3 m/s ) = 152.1 m/s m air = ρ1 Ac1V1 = (3.871 kg/m 3 )(0.1× 0.1 m 2 )(152.1 m/s) = 5.890 kg/s
The Rayleigh flow functions corresponding to the inlet and exit Mach numbers are T02/T0* = 1 (since Ma2 = 1) . T01 T0*
=
(k + 1)Ma 12 [2 + (k − 1)Ma 12 ] (1 + kMa 12 ) 2
=
(1.4 + 1)0.4 2 [2 + (1.4 − 1)0.4 2 ] (1 + 1.4 × 0.4 2 ) 2
= 0.5290
Therefore, T0 2 T02 / T0* 1 = = * T0 1 T01 / T0 0.5290
→ T02 = T01 / 0.5290 = (371.5 K ) / 0.5290 = 702.3 K
Then the rate of heat transfer becomes Q = m air c p (T02 − T01 ) = ( 5.890 kg/s )(1.005 kJ/kg ⋅ K )( 702.3 − 371.5 ) K = 1958 kW ≅ 1960 kW Discussion It can also be shown that T2 = 585 K, which is the highest thermodynamic temperature that can be attained under stated conditions. If more heat is transferred, the additional temperature rise will cause the mass flow rate to decrease. We can also solve this problem using the Rayleigh function values listed in Table A-15.
12-100 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-162 Solution Helium flowing at a subsonic velocity in a duct is accelerated by heating. The highest rate of heat transfer without affecting the inlet conditions is to be determined.
Assumptions 1 The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. 2 Inlet conditions (and thus the mass flow rate) remain constant. Properties
We take the properties of helium to be k = 1.667, cp = 5.193 kJ/kg⋅K, and R = 2.077 kJ/kg⋅K.
Analysis Heat transfer will stop when the flow is choked, and thus Ma2 = V2/c2 = 1. The inlet density and stagnation temperature are Q P1 400 kPa 3 = = 0.5350 kg/m ρ1 = RT1 (2.077 kJ/kgK)(360 K) P1 = 400 kPa T1 = 360 K Ma2 = 1 ⎛ k −1 ⎞ ⎛ 1.667 - 1 ⎞ T01 = T1 ⎜1 + Ma 12 ⎟ = (360 K)⎜1 + 0.4 2 ⎟ = 379.2 K 2 2 ⎝ ⎠ ⎝ ⎠ Ma1 = 0.4 Then the inlet velocity and the mass flow rate become ⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.667)(2.077 kJ/kg ⋅ K)(360 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 1116 m/s ⎟ ⎠
V1 = Ma 1c1 = 0.4(1116 m/s ) = 446.6 m/s
m air = ρ 1 Ac1V1 = (0.5350 kg/m 3 )(0.1× 0.1 m 2 )(446.6 m/s) = 2.389 kg/s
The Rayleigh flow functions corresponding to the inlet and exit Mach numbers are T02/T0* = 1 (since Ma2 = 1) T01 T0*
=
(k + 1)Ma 12 [2 + (k − 1)Ma 12 ] (1 + kMa 12 ) 2
=
(1.667 + 1)0.4 2 [2 + (1.667 − 1)0.4 2 ] (1 + 1.667 × 0.4 2 ) 2
= 0.5603
Therefore, T0 2 T02 / T0* 1 = = * T0 1 T01 / T0 0.5603
→ T02 = T01 / 0.5603 = (379.2 K ) / 0.5603 = 676.8 K
Then the rate of heat transfer becomes Q = m air c p (T02 − T01 ) = (2.389 kg/s)(5.193 kJ/kg ⋅ K )(676.8 − 379.2) K = 3690 kW
Discussion It can also be shown that T2 = 508 K, which is the highest thermodynamic temperature that can be attained under stated conditions. If more heat is transferred, the additional temperature rise will cause the mass flow rate to decrease. Also, in the solution of this problem, we cannot use the values of Table A-15 since they are based on k = 1.4.
12-101 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-163 Solution Air flowing at a subsonic velocity in a duct is accelerated by heating. For a specified exit Mach number, the heat transfer for a specified exit Mach number as well as the maximum heat transfer are to be determined.
Assumptions 1 The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. 2 Inlet conditions (and thus the mass flow rate) remain constant. Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K. Analysis
P1 = 35 kPa T1 = 400 K
Ma2 = 0.8
V1 = 100 m/s
The inlet Mach number and stagnation temperature are
⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(400 K)⎜⎜ ⎝ 1 kJ/kg Ma 1 =
q
⎞ ⎟ = 400.9 m/s ⎟ ⎠
V1 100 m/s = = 0.2494 c1 400.9 m/s
⎛ k −1 ⎞ ⎛ 1.4 - 1 ⎞ T01 = T1 ⎜1 + Ma 12 ⎟ = (400 K)⎜1 + 0.2494 2 ⎟ = 405.0 K 2 2 ⎝ ⎠ ⎝ ⎠
The Rayleigh flow functions corresponding to the inlet and exit Mach numbers are (Table A-15): Ma1 = 0.2494: Ma2 = 0.8:
T01/T* = 0.2559 T02/T* = 0.9639
Then the exit stagnation temperature and the heat transfer are determined to be T0 2 T02 / T * 0.9639 = = = 3.7667 T0 1 T01 / T * 0.2559
→ T0 2 = 3.7667T01 = 3.7667(405.0 K ) = 1526 K
q = c p (T02 − T01 ) = (1.005 kJ/kg ⋅ K )(1526 − 405 ) K = 1126 kJ/kg ≅ 1130 kJ/kg
Maximum heat transfer will occur when the flow is choked, and thus Ma2 = 1 and thus T02/T* = 1. Then, T0 2 T02 / T * 1 = = → T0 2 = T01 / 0.2559 = (405 K) / 0.2559 = 1583 K T0 1 T01 / T * 0.2559 qmax = c p ( T02 − T01 ) = ( 1.005 kJ/kg ⋅ K )( 1583 − 405 ) K = 1184 kJ/kg ≅ 1180 kJ/kg
Discussion This is the maximum heat that can be transferred to the gas without affecting the mass flow rate. If more heat is transferred, the additional temperature rise will cause the mass flow rate to decrease.
12-102 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow 12-164 Solution Air flowing at sonic conditions in a duct is accelerated by cooling. For a specified exit Mach number, the amount of heat transfer per unit mass is to be determined.
Assumptions The assumptions associated with Rayleigh flow (i.e., steady one-dimensional flow of an ideal gas with constant properties through a constant cross-sectional area duct with negligible frictional effects) are valid. Properties
We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K.
Analysis
Noting that Ma1 = 1, the inlet stagnation temperature is
⎛ k −1 ⎞ ⎛ 1.4 - 1 2 ⎞ T01 = T1 ⎜1 + Ma 12 ⎟ = (500 K)⎜1 + 1 ⎟ = 600 K 2 2 ⎝ ⎠ ⎝ ⎠
The Rayleigh flow functions T0/T0* corresponding to the inlet and exit Mach numbers are (Table A-15): Ma1 = 1: Ma2 = 1.6:
q P01 = 420 kPa T01 = 500 K
Ma2 = 1.6
Ma1 = 1
*
T01/T0 = 1 T02/T0* = 0.8842
Then the exit stagnation temperature and heat transfer are determined to be T0 2 T02 / T0* 0.8842 = = = 0.8842 T0 1 T01 / T0* 1
→ T02 = 0.8842T01 = 0.8842(600 K ) = 530.5 K
q = c p (T02 − T01 ) = (1.005 kJ/kg ⋅ K )(530.5 − 600) K = - 69.8 kJ/kg
Discussion The negative sign confirms that the gas needs to be cooled in order to be accelerated. Also, it can be shown that the thermodynamic temperature drops to 351 K at the exit
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Chapter 12 Compressible Flow 12-165 Solution Combustion gases enter a constant-area adiabatic duct at a specified state, and undergo a normal shock at a specified location. The exit velocity, temperature, and pressure are to be determined. Ma* = 1 Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant P1 = 180 kPa T* cross-sectional area adiabatic duct) are valid. 2 The friction factor is T1 = 510 K Normal P* constant along the duct. shock Ma1 = 2 V* Properties The specific heat ratio and gas constant of combustion gases are given to be k = 1.33 and R = 0.280 kJ/kg⋅K. The friction factor is given L1 = 2 m to be f = 0.010. Analysis The Fanno flow functions corresponding to the inlet Mach number of 2 are calculated from the relations in Table A-16 for k = 1.33 to be (fL*/Dh)1 = 0.3402 T1/T* = 0.7018, P1/P* = 0.4189 Ma1 = 2: First we check to make sure that the flow everywhere upstream the shock is supersonic. The required duct length from the D 0.10 m inlet L1* for the flow to reach sonic conditions is L*1 = 0.3402 = 0.3402 = 3.40 m , which is greater than the actual f 0.010 length of 2 m. Therefore, the flow is indeed supersonic when the normal shock occurs at the indicated location. Also, using fL1 (0.010)(2 m) the actual duct length L1, we have = = 0.2000 . Noting that L1 = L*1 − L*2 , the function fL*/Dh at the exit 0.10 m Dh ⎛ fL* ⎛ fL* ⎞ ⎟ =⎜ state and the corresponding Mach number are ⎜⎜ ⎟ ⎜ ⎝ Dh ⎠ 2 ⎝ Dh
From the relations in Table A-16, at Ma2 =1.476: velocity before the shock are determined to be
⎞ fL ⎟ − 1 = 0.3402 − 0.2000 = 0.1402 → Ma2 = 1.476. ⎟ ⎠1 Dh
T2/T* = 0.8568, P2/P* = 0.6270. Then the temperature, pressure, and
T2 T2 / T * 0.8568 = = = 1.2209 T1 T1 / T * 0.7018
→ T2 = 1.2209T1 = 1.2209(510 K ) = 622.7 K
P2 P2 / P * 0.6270 = = = 1.4968 P1 P1 / P * 0.4189
→ P2 = 1.4968 P1 = 1.4968(180 kPa ) = 269.4 kPa
The normal shock functions corresponding to a Mach number of 1.476 are, from the relations in Table A-14, Ma2 =1.476: Ma3 = 0.7052, T3/T2 = 1.2565, P3/P2 = 2.3466 Then the temperature and pressure after the shock become T3 = 1.2565T2 = 1.2565(622.7 K ) = 782.4 K P3 = 2.3466 P2 = 2.3466(269.4 kPa ) = 632.3 kPa
Sonic conditions exist at the duct exit, and the flow downstream of the shock is still Fanno flow. From the relations in Table A-16, Ma3 = 0.7052: T3/T* = 1.0767, P3/P* = 1.4713 Ma4 = 1: T4/T* = 1, P4/P* = 1 Then the temperature, pressure, and velocity at the duct exit are determined to be T4 T4 / T * 1 = = T3 T3 / T * 1.0767
→ T4 = T3 / 1.0767 = (782.4 K ) / 1.0767 = 727 K
P4 P4 / P * 1 = = * P3 P3 / P 1.4713
→ P4 = P3 / 1.4713 = (632.3 kPa ) / 1.4713 = 430 kPa
⎛ 1000 m 2 / s 2 V 4 = Ma 4 c 4 = (1) kRT4 = (1.33)(0.280 kJ/kg ⋅ K)(727 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 520 m/s ⎟ ⎠
Discussion It can be shown that L3* = 2.13 m, and thus the total length of this duct is 4.13 m. If the duct is extended, the normal shock will move farther upstream, and eventually to the inlet of the duct.
12-104 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-166
Solution Choked supersonic airflow in a constant cross-sectional area adiabatic duct is considered. The variation of duct length with Mach number is to be investigated, and the results are to be plotted. Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor remains constant along the duct. Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K. The average friction factor is given to be f = 0.03.
Ma2 = 1
Analysis We use EES to solve the problem. The flow is choked, and thus Ma2 =1. Corresponding to the inlet Mach number of Ma1 = 3 we have, from Table A-16, fL*/Dh = 0.5222, Therefore, the original duct length is D 0.18 m L*1 = 0.5222 = 0.5222 = 3.13 m f 0.03 Repeating the calculations for different Ma2 as it varies from 3 to 1 results in the following table for the location on the duct from the inlet:
P1 = 80 kPa T1 = 500 K Ma1 = 3
D =18 cm
T* P* V*
L1*
0
L
3
Duct length L, m 0.00 0.39 0.78 1.17 1.57 1.96 2.35 2.74 3.13
2.6
2.2
Mach number Ma
Mach number, Ma 3.00 2.75 2.50 2.25 2.00 1.75 1.50 1.25 1.00
1.8
1.4
1 0
EES program:
0.5
1
1.5
2
2.5
3
3.5
Duct length L, m
k=1.4 cp=1.005 R=0.287 P1=80 T1=500 Ma1=3 "Ma2=1" f=0.03 D=0.18 C1=sqrt(k*R*T1*1000) Ma1=V1/C1 T01=T02 T01=T1*(1+0.5*(k-1)*Ma1^2) T02=T2*(1+0.5*(k-1)*Ma2^2) P01=P1*(1+0.5*(k-1)*Ma1^2)^(k/(k-1)) rho1=P1/(R*T1) Ac=pi*D^2/4 mair=rho1*Ac*V1 P01Ps=((2+(k-1)*Ma1^2)/(k+1))^(0.5*(k+1)/(k-1))/Ma1
12-105 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow P1Ps=((k+1)/(2+(k-1)*Ma1^2))^0.5/Ma1 T1Ts=(k+1)/(2+(k-1)*Ma1^2) R1Rs=((2+(k-1)*Ma1^2)/(k+1))^0.5/Ma1 V1Vs=1/R1Rs fLs1=(1-Ma1^2)/(k*Ma1^2)+(k+1)/(2*k)*ln((k+1)*Ma1^2/(2+(k-1)*Ma1^2)) Ls1=fLs1*D/f P02Ps=((2+(k-1)*Ma2^2)/(k+1))^(0.5*(k+1)/(k-1))/Ma2 P2Ps=((k+1)/(2+(k-1)*Ma2^2))^0.5/Ma2 T2Ts=(k+1)/(2+(k-1)*Ma2^2) R2Rs=((2+(k-1)*Ma2^2)/(k+1))^0.5/Ma2 V2Vs=1/R2Rs fLs2=(1-Ma2^2)/(k*Ma2^2)+(k+1)/(2*k)*ln((k+1)*Ma2^2/(2+(k-1)*Ma2^2)) Ls2=fLs2*D/f L=Ls1-Ls2 P02=P02Ps/P01Ps*P01 P2=P2Ps/P1Ps*P1 V2=V2Vs/V1Vs*V1
Discussion
Note that the Mach number decreases nearly linearly along the duct.
12-106 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow
12-167
Solution Choked subsonic airflow in a constant cross-sectional area adiabatic duct is considered. The effect of duct length on the mass flow rate and the inlet conditions is to be investigated as the duct length is doubled. Assumptions 1 The assumptions associated with Fanno flow (i.e., steady, frictional flow of an ideal gas with constant properties through a constant cross-sectional area adiabatic duct) are valid. 2 The friction factor remains constant along the duct. Properties We take the properties of air to be k = 1.4, cp = 1.005 kJ/kg⋅K, and R = 0.287 kJ/kg⋅K. The average friction factor is given to be f = 0.02.
We use EES to solve the problem. The flow is choked, and thus Ma2 =1. The inlet Mach number is
Analysis
⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(400 K)⎜⎜ ⎝ 1 kJ/kg Ma 1 =
⎞ ⎟ = 400.9 m/s ⎟ ⎠
V1 120 m/s = = 0.2993 c1 400.9 m/s
Ma2 = 1 P1 = 100 kPa T1 = 400 K V1 = 120 m/s
Corresponding to this Mach number we have, from Table A16, fL*/Dh = 5.3312, Therefore, the original duct length is L = L*1 = 5.3312
D 0.06 m = 5.3312 = 16.0 m f 0.02
L1*
0
Then the initial mass flow rate becomes
ρ1 =
D = 6 cm
T* P* V*
L
P1 100 kPa = = 0.8711 kg/m 3 RT1 (0.287 kJ/kgK)(400 K)
m air = ρ 1 Ac1V1 = (0.8711 kg/m 3 )[π (0.06 m) 2 / 4](120 m/s) = 0.296 kg/s
Inlet velocity V1, m/s 129 126 123 120 117 115 112 110 108 106 104 103 101 99
0.32
Mass flow rate m air , kg/s 0.319 0.310 0.303 0.296 0.289 0.283 0.277 0.271 0.266 0.262 0.257 0.253 0.249 0.245
0.31 0.3 0.29
mair, kg/s
Duct length L, m 13 14 15 16 17 18 19 20 21 22 23 24 25 26
0.28 0.27 0.26 0.25 0.24 12
14
16
18
20
22
24
26
Duct lenght L, m
The EES program is listed below, along with a plot of inlet velocity vs. duct length:
12-107 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 12 Compressible Flow k=1.4 cp=1.005 R=0.287
C1=sqrt(k*R*T1*1000) Ma1=V1/C1 T01=T02 T01=T1*(1+0.5*(k-1)*Ma1^2) T02=T2*(1+0.5*(k-1)*Ma2^2) P01=P1*(1+0.5*(k-1)*Ma1^2)^(k/(k-1)) rho1=P1/(R*T1) Ac=pi*D^2/4 mair=rho1*Ac*V1
125 120 115
V1, m/s
P1=100 T1=400 "L=26" Ma2=1 f=0.02 D=0.06
130
110 105 100 95 12
14
16
18
20
22
24
26
Duct length L, m
P01Ps=((2+(k-1)*Ma1^2)/(k+1))^(0.5*(k+1)/(k-1))/Ma1 P1Ps=((k+1)/(2+(k-1)*Ma1^2))^0.5/Ma1 T1Ts=(k+1)/(2+(k-1)*Ma1^2) R1Rs=((2+(k-1)*Ma1^2)/(k+1))^0.5/Ma1 V1Vs=1/R1Rs fLs1=(1-Ma1^2)/(k*Ma1^2)+(k+1)/(2*k)*ln((k+1)*Ma1^2/(2+(k-1)*Ma1^2)) Ls1=fLs1*D/f P02Ps=((2+(k-1)*Ma2^2)/(k+1))^(0.5*(k+1)/(k-1))/Ma2 P2Ps=((k+1)/(2+(k-1)*Ma2^2))^0.5/Ma2 T2Ts=(k+1)/(2+(k-1)*Ma2^2) R2Rs=((2+(k-1)*Ma2^2)/(k+1))^0.5/Ma2 V2Vs=1/R2Rs fLs2=(1-Ma2^2)/(k*Ma2^2)+(k+1)/(2*k)*ln((k+1)*Ma2^2/(2+(k-1)*Ma2^2)) Ls2=fLs2*D/f L=Ls1-Ls2 P02=P02Ps/P01Ps*P01 P2=P2Ps/P1Ps*P1 V2=V2Vs/V1Vs*V1
Discussion Note that once the flow is choked, any increase in duct length results in a decrease in the mass flow rate and the inlet velocity.
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Chapter 12 Compressible Flow 12-168 Solution The flow velocity of air in a channel is to be measured using a Pitot-static probe, which causes a shock wave to occur. For measured values of static pressure before the shock and stagnation pressure and temperature after the shock, the flow velocity before the shock is to be determined.
Assumptions
1 Air is an ideal gas with constant specific heats. 2 Flow through the nozzle is steady and one-dimensional.
Properties
The specific heat ratio of air at room temperature is k= 1.4.
Analysis The nose of the probe is rounded (instead of being pointed), and thus it will cause a bow shock wave to form. Bow shocks are difficult to analyze. But they are normal to the body at the nose, and thus we can approximate them as normal shocks in the vicinity of the probe. It is given that the static pressure before the shock is P1 = 110 kPa, and the stagnation pressure and temperature after the shock are P02 = 620 kPa, and T02 = 340 K. Noting that the stagnation temperature remains constant, we have T01 = T02 = 340 K
Also,
P02 620 kPa = = 5.6364 ≈ 5.64 P1 110 kPa
P1 = 110 kPa
The fluid properties after the shock are related to those before the shock through the functions listed in Table A-14. For P02 / P1 = 5.64 we read Ma 1 = 2.0,
Ma 2 = 0.5774,
P02 = 0.7209, P01
P02 = 620 kPa T02 = 340 K
Shock wave
V1 ρ 2 = = 2.6667, V2 ρ1
Then the stagnation pressure and temperature before the shock become P01 = P02 /0.7209 = (620 kPa)/0.7209 = 860 kPa ⎛ P ⎞ T1 = T01 ⎜⎜ 1 ⎟⎟ ⎝ P01 ⎠
( k −1) / k
⎛ 110 kPa ⎞ = (340 K)⎜ ⎟ ⎝ 860 kPa ⎠
(1.4 −1) / 1.4
= 188.9 K
The flow velocity before the shock can be determined from V1 = Ma1c1, where c1 is the speed of sound before the shock, ⎛ 1000 m 2 / s 2 c1 = kRT1 = (1.4)(0.287 kJ/kg ⋅ K)(188.9 K)⎜⎜ ⎝ 1 kJ/kg
⎞ ⎟ = 275.5 m/s ⎟ ⎠
V1 = Ma 1c1 = 2(275.5 m/s) = 551 m/s
Discussion The flow velocity after the shock is V2 = V1/2.6667 = 551/2.6667 = 207 m/s. Therefore, the velocity measured by a Pitot-static probe would be very different that the flow velocity.
Design and Essay Problems
12-169 to 12-171 Solution Students’ essays and designs should be unique and will differ from each other.
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Chapter 13 Open-Channel Flow
Solutions Manual for Fluid Mechanics: Fundamentals and Applications by Çengel & Cimbala
CHAPTER 13 OPEN-CHANNEL FLOW
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Chapter 13 Open-Channel Flow Classification, Froude Number, and Wave Speed
13-1C Solution
We are to discuss differences between open-channel flow and internal flow.
Analysis Open-channel flow is the flow of liquids in channels open to the atmosphere or in partially filled conduits, and is characterized by the presence of a liquid-gas interface called the free surface, whereas internal flow is the flow of liquids or gases that completely fill a conduit. Discussion At a free surface of a liquid, the pressure must be equal to the pressure of the gas above it. This is what controls the pressure in open-channel flow. There is no such restriction in internal (pipe) flow since there is no free surface.
13-2C Solution
We are to discuss the driving force in open-channel flow and how flow rate is determined.
Analysis Flow in a channel is driven naturally by gravity. Water flow in a river, for example, is driven by the elevation difference between the source and the sink. The flow rate in an open channel is established by the dynamic balance between gravity and friction. Inertia of the flowing fluid also becomes important in unsteady flow. Discussion
In pipe flow, on the other hand, there may be an additional driving force of pressure due to pumps.
13-3C Solution
We are to discuss how pressure changes along the free surface in open-channel flow.
Analysis free surface.
The free surface coincides with the hydraulic grade line (HGL), and the pressure is constant along the
Discussion
At a free surface of a liquid, the pressure must be equal to the pressure of the gas above it.
13-4C Solution
We are to determine if the slope of the free surface is equal to the slope of the channel bottom.
Analysis No in general. The slope of the free surface is not necessarily equal to the slope of the bottom surface even during steady fully developed flow. Discussion
However, there are situations called uniform flow in which the conditions here are met.
13-5C Solution
We are to discuss the difference between uniform and nonuniform flow.
Analysis The flow in a channel is said to be uniform if the flow depth (and thus the average velocity) remains constant. Otherwise, the flow is said to be nonuniform or varied, indicating that the flow depth varies with distance in the flow direction. Uniform flow conditions are commonly encountered in practice in long straight sections of channels with constant slope and constant cross-section. Discussion
In uniform open-channel flow, the head loss due to frictional effects equals the elevation drop.
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Chapter 13 Open-Channel Flow 13-6C Solution
We are to define normal depth and how it is established.
Analysis In open channels of constant slope and constant cross-section, the fluid accelerates until the head loss due to frictional effects equals the elevation drop. The fluid at this point reaches its terminal velocity, and uniform flow is established. The flow remains uniform as long as the slope, cross-section, and the surface roughness of the channel remain unchanged. The flow depth in uniform flow is called the normal depth yn, which is an important characteristic parameter for open-channel flows. Discussion
The normal depth is a fairly strong function of surface roughness.
13-7C Solution We are to discuss some reasons for nonuniform flow in open channels, and the difference between rapidly varied flow and gradually varied flow. Analysis The presence of an obstruction in a channel such as a gate or a change in slope or cross-section causes the flow depth to vary, and thus the flow to become varied or nonuniform. The varied flow is called rapidly varied flow (RVF) if the flow depth changes markedly over a relatively short distance in the flow direction (such as the flow of water past a partially open gate or shortly before a falls), and gradually varied flow (GVF) if the flow depth changes gradually over a long distance along the channel. Discussion
The equations of GVF are simplified because of the slow changes in the flow direction.
13-8C Solution
We are to define and discuss hydraulic radius.
Analysis The hydraulic radius Rh is defined as the ratio of the cross-sectional flow area Ac and the wetted perimeter p. That is, Rh = Ac/p. Knowing the hydraulic radius, the hydraulic diameter is determined from Dh = 4Rh. Discussion
It is unfortunate (and not our fault!) that hydraulic radius is ¼ rather than ½ of hydraulic diameter.
13-9C Solution
We are to explain how to determine if a flow is tranquil, critical, or rapid.
Analysis
Knowing the average flow velocity and flow depth, the Froude number is determined from Fr = V / gy .
Then the flow is classified as Fr < 1 Fr = 1 Fr > 1
Subcritical or tranquil flow Critical flow Supercritical or rapid flow
Discussion
The Froude number is the most important parameter in open-channel flow.
13-10C Solution
We are to define and discuss the usefulness of the Froude number.
Analysis
Froude number, defined as Fr = V / gy , is a dimensionless parameter that governs the character of
flow in open channels. Here, g is the gravitational acceleration, V is the mean fluid velocity at a cross-section, and Lc is a characteristic length (Lc = flow depth y for wide rectangular channels). Fr represents the ratio of inertia forces to viscous forces in open-channel flow. The Froude number is also the ratio of the flow speed to wave speed, Fr = V /co. Discussion
The Froude number is the most important parameter in open-channel flow.
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Chapter 13 Open-Channel Flow 13-11C Solution
We are to define critical length, and discuss how it is determined.
Analysis The flow depth yc corresponding to a Froude number of Fr = 1 is the critical depth, and it is determined from V = gy c or y c = V 2 / g . Discussion
Critical depth is a useful parameter, even if the depth does not actually equal yc anywhere in the flow.
13-12C Solution We are to discuss whether the flow upstream of a hydraulic jump must be supercritical, and whether the flow downstream of a hydraulic jump must be subcritical. Analysis Upstream of a hydraulic jump, the upstream flow must be supercritical. Downstream of a hydraulic jump, the downstream flow must be subcritical. Discussion Otherwise, the second law of thermodynamics would be violated. Note that a hydraulic jump is analogous to a normal shock wave – in that case, the flow upstream must be supersonic and the flow downstream must be subsonic.
13-13 Solution The flow of water in a wide channel is considered. The speed of a small disturbance in flow for two different flow depths is to be determined for both water and oil. Assumptions negligible,
The distance across the wave is short and thus friction at the bottom surface and air drag at the top are
Analysis
Surface wave speed can be determined directly from the relation c 0 = gh .
(a) c 0 = gh = (9.81 m/s 2 ) (0.1 m) = 0.990 m/s (b) c 0 = gh = (9.81 m/s 2 ) (0.8 m) = 2.80 m/s Therefore, a disturbance in the flow will travel at a speed of 0.990 m/s in the first case, and 2.80 m/s in the second case. Discussion Note that wave speed depends on the water depth, and the wave speed increases as the water depth increases as long as the water remains shallow. Results would not change if the fluid were oil, because the wave speed depends only on the fluid depth.
13-14 Solution Water flows uniformly in a wide rectangular channel. For given flow depth and velocity, it is to be determined whether the flow is laminar or turbulent, and whether it is subcritical or supercritical. Assumptions Properties
The flow is uniform. The density and dynamic viscosity of water at 20ºC are ρ = 998.0 kg/m3 and μ = 1.002×10-3 kg/m⋅s.
Analysis
(a) The Reynolds number of the flow is Re =
3
ρVy (998.0 kg/m )(2 m/s )(0.2 m ) = = 3.984 × 10 5 , which is μ 1.002 × 10 − 3 kg/m ⋅ s
greater than the critical value of 500. Therefore, the flow is turbulent. 2 m/s V (b) The Froude number is Fr = = = 1.43 , which is greater than 1. gy (9.81 m/s 2 )(0.2 m) Therefore, the flow is supercritical. Discussion The result in (a) is expected since almost all open channel flows are turbulent. Also, hydraulic radius for a wide rectangular channel approaches the water depth y as the ratio y/b approaches zero.
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Chapter 13 Open-Channel Flow 13-15 Solution Water flow in a partially full circular channel is considered. For given water depth and average velocity, the hydraulic radius, Reynolds number, and the flow regime are to be determined. Assumptions
1 The flow is uniform.
Properties
The density and dynamic viscosity of water at 20ºC are ρ = 998.0 kg/m3 and μ = 1.002×10-3 kg/m⋅s.
From geometric considerations, R − a 1 − 0.5 2π π cos θ = = = 0.5 → θ = 60° = 60 = 1 R 360 3 R=1m Then the hydraulic radius becomes A π / 3 − sin(π / 3) cos(π / 3) θ − sin θ cos θ Rh = c = R= (1 m) = 0.293 m p 2θ 2π / 3 θ The Reynolds number of the flow is a=0.5 m ρVR h (998.0 kg/m 3 )(2 m/s )(0.293 m ) 5 = = 5.84 × 10 Re = μ 1.002 × 10 − 3 kg/m ⋅ s which is greater than the critical value of 500. Therefore, the flow is turbulent. When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, Ac = R 2 (θ − sin θ cos θ ) = (1 m) 2 [π / 3 − sin(π / 3) cos(π / 3)] = 0.6142 m 2 Analysis
yh =
0.6142 m 2 Ac Ac = = = 0.3546 m Top width 2 R sin θ 2(1 m)sin60°
→
Fr =
V gyh
=
2 m/s (9.81m/s 2 )(0.3546 m)
= 1.07
which is greater than 1. Therefore, the flow is supercritical (although, very close to critical). Discussion Note that if the maximum flow depth were used instead of the hydraulic depth, the result would be subcritical flow, which is not true.
13-16 Solution Water flows uniformly in a wide rectangular channel. For given values of flow depth and velocity, it is to be determined whether the flow is subcritical or supercritical. Assumptions
1 The flow is uniform. 2 The channel is wide and thus the side wall effects are negligible.
Analysis
The Froude number is Fr =
V gy
=
4 m/s (9.81 m/s 2 )(0.08 m)
= 4.51 , which is greater than 1.
Therefore, the flow is supercritical. Discussion temperature.
Note that the Froude Number is not function of any temperature-dependent properties, and thus
13-17 Solution Rain water flows on a concrete surface. For given values of flow depth and velocity, it is to be determined whether the flow is subcritical or supercritical. Assumptions
1 The flow is uniform. 2 The thickness of water layer is constant.
Analysis
The Froude number is Fr =
V gy
=
1.3 m/s (9.81 m/s 2 )(0.02 m)
= 2.93 , which is greater than 1.
Therefore, the flow is supercritical. Discussion
This water layer will undergo a hydraulic jump when the ground slope decreases or becomes adverse.
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Chapter 13 Open-Channel Flow
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Chapter 13 Open-Channel Flow 13-18E Solution Water flows uniformly in a wide rectangular channel. For given flow depth and velocity, it is to be determined whether the flow is laminar or turbulent, and whether it is subcritical or supercritical. Assumptions
The flow is uniform.
Properties
The density and dynamic viscosity of water at 70ºF are ρ = 62.30 lbm/ft3 and μ = 6.556×10-4 lbm/ft⋅s.
Analysis
(a) The Reynolds number of the flow is Re =
3
ρVy (62.30 lbm/ft )(6 ft/s)(0.5 ft ) = = 2.85 × 10 5 , which is −4 μ 6.556 × 10 lbm/ft ⋅ s
greater than the critical value of 500. Therefore, the flow is turbulent. (b) The Froude number is Fr =
V gy
=
6 ft/s (32.2 ft/s 2 )(0.5 ft)
= 1.50 , which is greater than 1.
Therefore, the flow is supercritical. Discussion The result in (a) is expected since almost all open channel flows are turbulent. Also, hydraulic radius for a wide rectangular channel approaches the water depth y as the ratio y/b approaches zero.
13-19 Solution Water flows uniformly through a half-full circular channel. For a given average velocity, the hydraulic radius, the Reynolds number, and the flow regime are to be determined. Assumptions
The flow is uniform.
Properties
The density and dynamic viscosity of water at 10ºC are ρ = 999.7 kg/m3 and μ = 1.307×10-3 kg/m⋅s.
Analysis
From geometric considerations, the hydraulic radius is
Rh =
Ac πR 2 / 2 R 1.5 m = = = = 0.75 m 2 2 p πR
R = 1.5 m
The Reynolds number of the flow is ρVR h (999.7 kg/m 3 )(2.5 m/s )(0.75 m ) = = 1.43 × 10 6 , which is Re = μ 1.307 × 10 − 3 kg/m ⋅ s greater than the critical value of 500. Therefore, the flow is turbulent. When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a nonrectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, Ac πR 2 / 2 πR π (1.5 m) yh = = = = = 1.178 m Top width 2R 4 4 Fr =
V gyh
=
2.5 m/s (9.81m/s 2 )(1.178 m)
θ = π/2
= 0.735 , which is greater than 1. Therefore, the flow is subcritical.
Discussion If the maximum flow depth were used instead of the hydraulic depth, the result would still be subcritical flow, but this is not always the case.
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Chapter 13 Open-Channel Flow 13-20 Solution A single wave is initiated in a sea by a strong jolt during an earthquake. The speed of the resulting wave is to be determined. Assumptions
The depth of water is constant,
Analysis
Surface wave speed is determined the wave-speed relation to be c 0 = gh = (9.81 m/s 2 ) (2000 m ) = 140 m/s
Discussion Note that wave speed depends on the water depth, and the wave speed increases as the water depth increases. Also, the waves eventually die out because of the viscous effects.
Specific Energy and the Energy Equation
13-21C Solution
We are to define and discuss specific energy.
The specific energy Es of a fluid flowing in an open channel is the sum of the pressure and dynamic V2 . heads of a fluid, and is expressed as E s = y + 2g
Analysis
Discussion
Specific energy is very useful when analyzing varied flows.
13-22C Solution
We are to compare the specific energy in two flows – one subcritical and one supercritical.
Analysis A plot of Es versus y for constant V through a rectangular channel of width b reveals that there are two y values corresponding to a fixed value of Es: one for subcritical flow and one for supercritical flow. Therefore, the specific energies of water in those two channels can be identical. Discussion
If the flow is varied (not uniform), however, Es is not necessarily identical in the two channels.
13-23C Solution
We are to examine claims about the minimum value of specific energy.
Analysis
The point of minimum specific energy is the critical point, and thus the first person is correct.
Discussion The specific energy cannot go below the critical point for a given volume flow rate, as is clear from the plot of specific energy as a function of flow depth.
13-24C Solution We are to examine a claim about supercritical flow of water in an open channel, namely, that the larger the flow depth, the larger the specific energy. Analysis No, the claim is incorrect. A plot of Es versus y for constant V reveals that the specific energy decreases as the flow depth increases during supercritical channel flow. Discussion This may go against our intuition, since a larger flow depth seems to imply greater energy, but this is not necessarily the case (we cannot always trust our intuition).
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Chapter 13 Open-Channel Flow 13-25C Solution
We are to examine a claim that specific energy remains constant in steady uniform flow.
Analysis The first person (who claims that specific energy remains constant) is correct since in uniform flow, the flow depth and the flow velocity, and thus the specific energy, remain constant since E s = y + V 2 / 2 g . The head loss is made up by the decline in elevation (the channel is sloped downward in the flow direction). Discussion In uniform flow, the flow depth and the average velocity do not change downstream, since the elevation drop exactly overcomes the frictional losses.
13-26C Solution
We are to define and discuss friction slope.
Analysis
The friction slope is related to head loss hL, and is defined as S f = h L / L where L is the channel length.
The friction slope is equal to the bottom slope when the head loss is equal to the elevation drop. That is, S f = S 0 when h L = z1 − z 2 . Discussion
Friction slope is a useful concept when analyzing uniform or varied flow in open channels.
13-27C Solution We are to examine a claim that during steady flow in a wide rectangular channel, the energy line of the flow is parallel to the channel bottom when the frictional losses are negligible. Analysis No, the claim is not correct. The energy line is a distance E s = y + V 2 / 2 g (total mechanical energy of the fluid) above a horizontal reference datum. When there is no head loss, the energy line is horizontal even when the channel is not. The elevation and velocity heads ( z + y and V 2 / 2 g ) may convert to each other during flow in this case, but their sum remains constant. Discussion Keep in mind that in real life, there is no such thing as frictionless flow. However, there are situations in which the frictional effects are negligible compared to other effects in the flow.
13-28C Solution We are to examine a claim that during steady 1-D flow through a wide rectangular channel, the total mechanical energy of the fluid at the free surface is equal to that of the fluid at the channel bottom. Analysis Yes, the claim is correct. During steady one-dimensional flow, the total mechanical energy of a fluid at any point of a cross-section is given by H = z + y + V 2 / 2 g . Discussion
The physical elevation of the point under consideration does not appear in the above equation for H.
13-29C Solution
We are to express the total mechanical energy in steady 1-D flow in terms of heads.
Analysis The total mechanical energy of a fluid at any point of a cross-section is expressed as H = z + y + V 2 / 2 g where y is the flow depth, z is the elevation of the channel bottom, and V is the average flow velocity. It is related to the specific energy of the fluid by H = z + E s . Discussion
Because of irreversible frictional head losses, H must decrease in the flow direction in open-channel flow.
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Chapter 13 Open-Channel Flow 13-30C Solution
We are to express the 1-D energy equation for open-channel flow and discuss head loss.
The one-dimensional energy equation for open channel flow between an upstream section 1 and V2 V2 downstream section 2 is written as z1 + y1 + 1 = z 2 + y 2 + 2 + h L where y is the flow depth, z is the elevation of the 2g 2g
Analysis
channel bottom, and V is the average flow velocity. The head loss hL due to frictional effects can be determined from L V2 hL = f where f is the average friction factor and L is the length of channel between sections 1 and 2. Rh 8g Discussion Head loss is always positive – it can never be negative since this would violate the second law of thermodynamics. Thus, the total mechanical energy must decrease downstream in open-channel flow.
13-31 Solution Water flow in a rectangular channel is considered. The character of flow, the flow velocity, and the alternate depth are to be determined. Assumptions
The specific energy is constant.
Analysis
The average flow velocity is determined from
V=
0.7 m3/s
0.7 m 3 /s V V = = = 3.50 m/s Ac yb (0.25 m)(0.8 m)
0.25 m
The critical depth for this flow is ⎛ V 2 yc = ⎜ 2 ⎜ gb ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (0.7 m 3 /s) 2 =⎜ ⎜ (9.81 m/s 2 )( 0.8 m ) 2 ⎝
⎞ ⎟ ⎟ ⎠
0.8 m
1/ 3
= 0.427 m
Therefore, the flow is supercritical since the actual flow depth is y = 0.25 m, and y < yc. The specific energy for given conditions is E s1 = y1 +
V 2
2 gb 2 y12
= (0.25 m) +
(0.7 m 3 /s) 2 2(9.81 m/s 2 )(0.8 m) 2 (0.25 m) 2
= 0.874 m
Then the alternate depth is determined from E s1 = E s 2 to be (0.7 m 3 /s) 2 V 2 E s2 = y2 + 0 . 874 m y → = + 2 2 gb 2 y 22 2(9.81 m/s 2 )(0.8 m) 2 y 22 Solving for y2 gives the alternate depth to be y2 = 0.82 m. There are three roots of this equation; one for subcritical, one for supercritical and third one as a negative root. Therefore, if the character of flow is changed from supercritical to subcritical while holding the specific energy constant, the flow depth will rise from 0.25 m to 0.82 m. Discussion Two alternate depths show two possible flow conditions for a given specific energy. If the energy is not the minimum specific energy, there are two water depths corresponding to subcritical and supercritical states of flow. As an example, these two depths may be observed before and after a sluice gate as alternate depths, if the losses are disregarded.
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Chapter 13 Open-Channel Flow 13-32 Solution Water flows in a rectangular channel. The specific energy and whether the flow is subcritical or supercritical are to be determined. Assumptions
The flow is uniform and thus the specific energy is constant.
Analysis For convenience, we take the channel width to be b = 1 m. Then the volume flow rate and the critical depth for this flow become V = VAc = Vyb = (6 m/s)(0.4 m)(1 m) = 2.40 m 3 /s
⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (2.40 m 3 /s) 2 =⎜ ⎜ (9.81 m/s 2 )(1 m) 2 ⎝
⎞ ⎟ ⎟ ⎠
6 m/s
1/3
= 0.837 m
The flow is supercritical since the actual flow depth is y = 0.4 m, and y < yc. The specific energy for given conditions is E s1 = y1 +
Discussion
V 2 2 gb 2 y12
= y1 +
0.40 m
(6 m/s) 2 V2 = (0.4 m) + = 2.23 m 2g 2(9.81 m/s 2 )
Note that the flow may also exist as subcritical flow at the same value of specific energy,
13-33 [Also solved using EES on enclosed DVD] Solution Water flows in a rectangular channel. The critical depth, the alternate depth, and the minimum specific energy are to be determined. Assumptions
The channel is sufficiently wide so that the edge effects are negligible.
Analysis For convenience, we take the channel width to be b = 1 m. Then the volume flow rate and the critical depth for this flow become V = VAc = Vyb = (6 m/s)(0.4 m)(1 m) = 2.40 m 3 /s
⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (2.40 m 3 /s) 2 =⎜ ⎜ (9.81 m/s 2 )(1 m) 2 ⎝
⎞ ⎟ ⎟ ⎠
1/3
= 0.837 m
6 m/s
(b) The flow is supercritical since the actual flow depth is y = 0.4 m, and y < yc. The specific energy for given conditions is E s1 = y1 +
V 2 2 gb 2 y12
= y1 +
(6 m/s) 2 V2 = (0.4 m) + = 2.23 m 2g 2(9.81 m/s 2 )
0.40 m
Then the alternate depth is determined from E s1 = E s 2 to be 0.240 m 3 /s V 2 E s2 = y 2 + y 2.23 m → = + 2 2 gb 2 y 22 y 22 Solving for y2 gives the alternate depth to be y2 = 2.17 m. Therefore, if the character of flow is changed from supercritical to subcritical while holding the specific energy constant, the flow depth will rise from 0.4 m to 2.17 m. (c) the minimum specific energy is V2 gy 3 3 E s , min = y c + c = y c + c = y c = (0.837 m ) = 1.26 m 2g 2g 2 2 Discussion Note that minimum specific energy is observed when the flow depth is critical.
13-11
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Chapter 13 Open-Channel Flow 13-34
Solution Water flows in a rectangular channel. The critical depth, the alternate depth, and whether the flow is subcritical or supercritical are to be determined. Assumptions Analysis
The flow is uniform and thus the specific energy is constant. ⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
(a) The critical depth is calculated to be
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (12 m 3 /s) 2 = ⎜⎜ 2 2 ⎝ (9.81 m/s )(6 m)
⎞ ⎟ ⎟ ⎠
1/3
= 0.742 m
(b) The average flow velocity and the Froude number are 12 m 3 /s V V 3.636 m/s = = 1.565 , which is greater than 1. Fr1 = V= = = 3.636 m/s and b y (6 m) ( 0.55 m) gy (9.81 m/s 2 ) (0.55 m) Therefore, the flow is supercritical.
12 m3/s
(c) Specific energy for this flow is V 2
E s1 = y1 +
2 gb 2 y12
= (0.55 m) +
(12 m 3 /s) 2 2(9.81 m/s 2 )(6 m) 2 (0.55 m) 2
= 1.224 m
0.55 m Then the alternate depth is determined from Es1=Es2, E s2 = y2 +
V 2
6m
(12 m 3 /s) 2
→ 1.224 m = y 2 +
2 gb 2 y 22 2(9.81 m/s 2 )(6 m) 2 y 22 The alternate depth is calculated to be y2 = 1.03 m which is the subcritical depth for the same value of specific energy.
Discussion The depths 0.55 m and 1.03 are alternate depths for the given discharge and specific energy. The flow conditions determine which one is observed.
13-35E Solution Water flows in a wide rectangular channel. For specified values of flow depth and average velocity, the Froude number, critical depth, and whether the flow is subcritical or supercritical are to be determined. Assumptions
The flow is uniform and thus the specific energy is constant.
Analysis
(a) The Froude number is
Fr =
V gy
⎛ V 2 (b) The critical depth is calculated to be y c = ⎜ 2 ⎜ gb ⎝ (c) The flow is supercritical since Fr > 1.
⎞ ⎟ ⎟ ⎠
= 1/ 3
14 ft/s (32.2 ft/s 2 )(0.8 ft) ⎛ V 2 y 2b 2 =⎜ ⎜ gb 2 ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
= 2.76
⎛ (14 ft s) 2 (0.8 ft )2 =⎜ ⎜ (32.2 ft/s 2 ) ⎝
⎞ ⎟ ⎟ ⎠
1/3
14 ft/s
For the case of y = 0.2 ft: Replacing 0.8 ft in above calculations by 0.2 ft gives 14 ft/s V Fr = = = 5.52 gy (32.2 ft/s 2 )(0.2 ft) 1/ 3
1/ 3
2 2 ⎛ ⎛ V 2 ⎞ ⎛ V 2 y 2b 2 ⎞ ⎟ = ⎜ (14 ft s ) (0.2 ft ) yc = ⎜ 2 ⎟ = ⎜ ⎜ gb ⎟ ⎜ gb 2 ⎟ ⎜ 32.2 ft/s 2 ⎝ ⎠ ⎝ ⎠ ⎝ The flow is supercritical in this case also since Fr > 1.
(
Discussion
= 1.57 ft
)
⎞ ⎟ ⎟ ⎠
0.8 ft
1/3
= 0.625 ft
Note that the value of critical depth depends on flow rate, and it decreases as the flow rate decreases.
13-12
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-36E Solution Water flows in a wide rectangular channel. For specified values of flow depth and average velocity, the Froude number, critical depth, and whether the flow is subcritical or supercritical are to be determined. Assumptions
The flow is uniform and thus the specific energy is constant.
Analysis
(a) The Froude number is
Fr =
V gy
⎛ V 2 (b) The critical depth is calculated to be y c = ⎜ 2 ⎜ gb ⎝ (c) The flow is supercritical since Fr > 1.
⎞ ⎟ ⎟ ⎠
= 1/ 3
10 ft/s (32.2 ft/s 2 )(0.8 ft) ⎛ V 2 y 2b 2 =⎜ ⎜ gb 2 ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
= 1.97
⎛ (10 ft s) 2 (0.8 ft )2 =⎜ ⎜ (32.2 ft/s 2 ) ⎝
⎞ ⎟ ⎟ ⎠
1/3
10 ft/s
For the case of y = 0.2 ft: Replacing 0.8 ft in above calculations by 0.2 ft gives 10 ft/s V Fr = = = 3.94 gy (32.2 ft/s 2 )(0.2 ft) 1/ 3
1/ 3
2 2 ⎛ ⎛ V 2 ⎞ ⎛ V 2 y 2b 2 ⎞ ⎟ = ⎜ (14 ft s) (0.2 ft ) yc = ⎜ 2 ⎟ = ⎜ ⎜ gb ⎟ ⎜ gb 2 ⎟ ⎜ (32.2 ft/s 2 ) ⎝ ⎠ ⎝ ⎠ ⎝ The flow is supercritical in this case also since Fr > 1.
Discussion
= 1.26 ft
⎞ ⎟ ⎟ ⎠
0.8 ft
1/3
= 0.50 ft
Note that the value of critical depth depends on flow rate, and it decreases as the flow rate decreases.
13-37 Solution Critical flow of water in a rectangular channel is considered. For a specified average velocity, the flow rate of water is to be determined. Assumptions constant.
The flow is uniform and thus the specific energy is
5 m/s
Analysis The Froude number must be unity since the flow is critical, and thus Fr = V / gy = 1 . Therefore, y = yc =
(5 m /s) 2 V2 = = 2.55 m g 9.81 m/s 2
yc 4m
Then the flow rate becomes V = VAc = Vby = (5 m/s) (4 m ) (2.55 m) = 51.0 m 3 /s
Discussion Critical flow is not a stable type of flow and can be observed for short intervals. Occurrence of critical depth is important as boundary condition most of the time. For example it can be used as a flow rate computation mechanism for a channel ending with a drawdown.
13-13
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Chapter 13 Open-Channel Flow 13-38 Solution Water flows uniformly through a half-full circular steel channel. For a given average velocity, the volume flow rate, critical slope, and the critical depth are to be determined. Assumptions
The flow is uniform.
Analysis
The volume flow rate is determined from
π (0.25 m) 2 πR 2 V = VAc = V = (2.8 m/s) = 0.275 m 3 /s 2 2 When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, Ac πR 2 / 2 πR π (0.25 m) = = = = 0.1963 m yh = Top width 2R 4 4 2.8 m/s V Fr = = = 2.02 gy (9.81 m/s 2 )(0.1963 m)
R =0.25 m
θ = π/2
which is greater than 1. Therefore, the flow is supercritical. Note that if the maximum flow depth were used instead of the hydraulic depth, the result could be Discussion different, especially when the Froude number is close to 1.
13-39 Solution Water flows uniformly through a half-full hexagon channel. For a given flow rate, the average velocity and whether the flow is subcritical or supercritical are to be determined. Assumptions
The flow is uniform.
Analysis
(a) The flow area is determined from geometric considerations to be
( 2 + 2 × 2) m 2 m (b + 2b ) b tan 60° = 5.196 m 2 tan 60° = 2 2 2 2 Then the average velocity becomes 45 m 3 s V V= = = 8.66 m s Ac 5.196 m 2 (b) When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, 5.196 m Ac A = c = = 1.299 m y = yh = Top width 2b 2× 2 m Then the Froude number becomes 8.66 m/s V = = 2.43 Fr = gy (9.81 m/s 2 )(1.299 m) Ac =
h = (b/2)tan60 b=2m
which is greater than 1. Therefore, the flow is supercritical. Discussion The analysis is approximate since the edge effects are significant here compared to a wide rectangular channel, and thus the results should be interpreted accordingly.
13-14
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-40 Solution Water flows uniformly through a half-full hexagon channel. For a given flow rate, the average velocity and whether the flow is subcritical or supercritical are to be determined. Assumptions
The flow is uniform.
Analysis
The flow area is determined from geometric considerations to be
( 2 + 2 × 2) m 2 m (b + 2b ) b tan 60° = 5.196 m 2 tan 60° = 2 2 2 2 Then the average velocity becomes 30 m 3 s V V= = = 5.77 m s Ac 5.196 m 2 When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, 5.196 m 2 Ac A = c = = 1.299 m y = yh = Top width 2b 2× 2 m Then the Froude number becomes 5.77 m/s V = = 1.62 Fr = gy (9.81 m/s 2 )(1.299 m) Ac =
h = (b/2)tan60 b=2m
which is greater than 1. Therefore, the flow is supercritical. Discussion The analysis is approximate since the edge effects are significant here compared to a wide rectangular channel, and thus the results should be interpreted accordingly.
Uniform Flow and Best Hydraulic Cross Sections
13-41C Solution
We are to discuss when flow in an open channel is uniform, and how it remains uniform.
Analysis Flow in a channel is called uniform flow if the flow depth (and thus the average flow velocity) remains constant. The flow remains uniform as long as the slope, cross-section, and the surface roughness of the channel remain unchanged. Discussion
Uniform flow in open-channel flow is somewhat analogous to fully developed pipe flow in internal flow.
13-42C Solution
We are to discuss how flow depth changes when the bottom slope is increased.
Analysis
The flow depth decreases when the bottom slope is increased.
Discussion You can think of it in simple terms this way: As the slope increases, the liquid flows faster, and faster flow requires lower depth.
13-15
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Chapter 13 Open-Channel Flow 13-43C Solution
We are to examine a claim that head loss can be determined by multiplying bottom slope by channel length.
Analysis Yes, the claim is correct. The head loss in uniform flow is h L = S 0 L since the head loss must equal elevation loss. Discussion In uniform flow, frictional head losses are exactly balanced by elevation loss, which is directly proportional to bottom slope.
13-44C Solution
We are to discuss the constants and coefficients in the Manning equation.
Analysis
The value of the factor a in SI units is a = 1 m1/3/s. Combining the relations C = 8 g / f and C =
and solving them for n gives the desired relation to be n =
a 8g / f
a 1/ 6 Rh n
R h1 / 6 . In practice, n is usually determined
experimentally. Discussion
The value of n varies greatly with surface roughness.
13-45C
Solution Sc =
gn 2 a 2 y c1 / 3
It is to be shown that for uniform critical flow, the general critical slope relation S c =
gn 2 y c a 2 R h4 / 3
reduces to
for film flow with b >> yc.
For critical flow, the flow depth is y = yc. For film flow, the hydraulic radius is Rh = y = yc. Substituting into gn 2 y gn 2 y gn 2 the critical slope relation gives the desired result, S c = 2 4 /c3 = 2 4 /c3 = 2 1 / 3 . a Rh a yc a yc
Analysis
Discussion
The reduced equation is valid for film flow only – be careful not to apply it to channels of other shapes.
13-46C Solution
We are to determine which cross section is better – one with a small or large hydraulic radius.
Analysis The best hydraulic cross-section for an open channel is the one with the maximum hydraulic radius, or equivalently, the one with the minimum wetted perimeter for a specified cross-sectional area. Discussion Frictional losses occur at the wetted perimeter walls of the channel, so it makes sense to minimize the wetted perimeter in order to minimize the frictional losses.
13-47C Solution
We are to determine which cross section shape is best for an open channel.
Analysis
The best hydraulic cross-section for an open channel is a (a) circular one.
Discussion
Circular channels are often more difficult to construct, however, so they are often not used in practice.
13-16
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Chapter 13 Open-Channel Flow 13-48C Solution
We are to determine the best hydraulic cross section for a rectangular channel.
Analysis width.
The best hydraulic cross section for a rectangular channel is one whose fluid height is (a) half the channel
Discussion
It turns out that for this case, the wetted perimeter, and thus the frictional losses, are smallest.
13-49C Solution
We are to determine the best hydraulic cross section for a trapezoidal channel.
Analysis The best hydraulic cross section for a trapezoidal channel of base width b is (a) one for which the length of the side edge of the flow section is b. Discussion
It turns out that for this case, the wetted perimeter, and thus the frictional losses, are smallest.
13-50C Solution
We are to determine how the flow rate changes when the Manning coefficient doubles.
a The flow rate in uniform flow is given as V = Ac R h2 / 3 S 01 / 2 , and thus the flow rate is inversely n proportional to the Manning coefficient. Therefore, if the Manning coefficient doubles as a result of some algae growth on surfaces while the flow cross section remains constant, the flow rate will (d) decrease by half.
Analysis
In an actual case, the cross section may also change due to flow depth changes as well.
Discussion
13-51 Solution The flow of water in a trapezoidal finished-concrete channel is considered. For a given flow depth and bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel of finished concrete is n = 0.012 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are 0.45 m ⎞ y ⎞ ⎛ ⎛ Ac = y⎜ b + ⎟ = 0.3724 m 2 ⎟ = (0.45 m) ⎜ 0.60 m + tan θ ⎠ tan50° ⎠ ⎝ ⎝
Analysis
p =b+ Rh =
2 (0.45 m) 2y = 0.6 m + = 1.775 m sin θ sin50°
y = 0.45 m
Ac 0.3724 m 2 = = 0.2096 m p 1.775 m
θ =50° b = 0.6 m
Bottom slope of the channel is S 0 = tan 0.4 D = 0.006981
Then the flow rate can be determined from Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (0.3724 m 2 )(0.2096 m) 2/3 (0.006981) 1 / 2 = 0.915 m 3 s n 0.012
Discussion
Note that the flow rate in a given channel is a strong function of the bottom slope.
13-17
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Chapter 13 Open-Channel Flow 13-52 Solution Water flows uniformly half-full in a circular finished-concrete channel. For a given bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel of finished concrete is n = 0.012 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are π (1 m) 2 πR Ac = = = 1.571 m 2 2 2 2πR 2π (1 m) = = 3.142 m p= 2 2 A πR 2 / 2 R 1 m = = = 0.50 m Rh = c = 2 2 πR P
Analysis
2
R=1m
θ = π/2
Then the flow rate can be determined from Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (1.571 m 2 )(0.50 m) 2/3 (1.5 / 1000) 1 / 2 = 3.19 m 3 s n 0.012
Discussion
Note that the flow rate in a given channel is a strong function of the bottom slope.
13-53E Solution Water is to be transported uniformly in a full semi-circular unfinished-concrete channel. For a specified flow rate, the elevation difference across the channel is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel of unfinished concrete is n = 0.014 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are π (3 ft) 2 Ac = = = 14.14 ft 2 2 2 2πR 2π (3 ft ) = = 9.425 ft p= 2 2 A πR 2 / 2 R 3 ft Rh = c = = = = 1.50 ft P 2 2 πR
Analysis
πR 2
R = 3 ft
θ = π/2
Substituting the given quantities into Manning’s equation, 1.486 ft 1/3 / s a (14.14 ft 2 )(1.50 ft) 2/3 S 01 / 2 V = Ac R h2 / 3 S 01 / 2 → 150 ft 3 /s = 0.014 n It gives the slope to be S0 = 0.005817. Therefore, the elevation difference Δz across a pipe length of L = 1 mile = 5280 ft must be Δz = S 0 L = 0.005817(5280 ft) = 30.7 ft
Discussion Note that when transporting water through a region of fixed elevation drop, the only way to increase the flow rate is to use a channel with a larger cross-section.
13-18
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Chapter 13 Open-Channel Flow 13-54 Solution Water is to be transported uniformly in a trapezoidal asphalt-lined channel. For a specified flow rate, the required elevation drop per km channel length is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an asphalt-lined open channel is n = 0.016 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are 10 m + 5 m Ac = (2.2 m) = 16.5 m 2 2
Analysis
p = (5 m) + 2 (2.2 m) 2 + (2.5 m) 2 = 11.66 m
10 m
2.2 m
A 16.5 m 2 Rh = c = = 1.415 m p 11.66 m
5m
Substituting the given quantities into Manning’s equation, a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 → 120 m 3 /s = (16.5 m 2 )(1.415 m) 2/3 S 01 / 2 n 0.016 It gives the slope to be S0 = 0.008524. Therefore, the elevation drop Δz across a pipe length of L = 1 km must be Δz = S 0 L = 0.008524(1000 m) = 8.52 m
Discussion Note that when transporting water through a region of fixed elevation drop, the only way to increase the flow rate is to use a channel with a larger cross-section.
13-19
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Chapter 13 Open-Channel Flow 13-55 Solution The flow of water through the trapezoidal asphalt-lined channel in the previous problem is reconsidered. The maximum flow rate corresponding to a given maximum channel height is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
We denote the flow conditions in the previous problem by subscript 1 and the conditions for the maximum a case in this problem by subscript 2. Using the Manning’s equation V = Ac R h2 / 3 S 01 / 2 and noting that the Manning n coefficient and the channel slope remain constant, the flow rate in case 2 can be expressed in terms of flow rate in case 1 as Analysis
V2 (a / n ) Ac 2 R h22/ 3 = V1 (a / n ) Ac1 R h21/ 3
A ⎛R → V2 = c 2 ⎜ h 2 Ac1 ⎜ R h1 ⎝
⎞ ⎟ ⎟ ⎠
2/3
10 m
V1
The trapezoid angle is tan θ = 2.2 / 2.5 = 0.88 → θ = 2.2 / 2.5 = 41.34° .
2.2 m
From geometric considerations, 10 m + 5 m Ac1 = (2.2 m) = 16.5 m 2 2
θ 5m
p1 = (5 m) + 2 (2.2 m) 2 + (2.5 m) 2 = 11.66 m
R h1 =
5+2×2.4/tanθ =10.45 m
Ac1 16.5 m 2 = = 1.415 m p1 11.66 m
and Ac 2 =
10.45 m + 5 m 2
2.4 m
(2.4 m) = 18.54 m 2
p 2 = (5 m) + 2 (2.4 m) 2 + (5.45/2 m) 2 = 12.26 m Rh2
5m
A 18.54 m 2 = c2 = = 1.512 m p2 12.26 m
Substituting, A ⎛R V2 = c 2 ⎜ h 2 Ac1 ⎜ R h1 ⎝
Discussion
⎞ ⎟ ⎟ ⎠
2/3
18.54 m 2 ⎛ 1.512 m ⎞ V1 = ⎟ ⎜ 16.5 m 2 ⎝ 1.415 m ⎠
2/3
(120 m 3 /s ) = 141 m 3 /s
Note that a 9% increase in flow depth results in an 18% increase in flow rate.
13-20
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-56 Solution The flow of water through two identical channels with square flow sections is considered. The percent increase in flow rate as a result of combining the two channels while the flow depth remains constant is to be determined. Assumptions 1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the channel. Analysis We denote the flow conditions for two separate channels by subscript 1 and the conditions for the combined wide channel by subscript 2. Using the Manning’s equation a V = Ac R h2 / 3 S 01 / 2 and noting that the Manning coefficient, channel n slope, and the flow area Ac remain constant, the flow rate in case 2 can be expressed in terms of flow rate in case 1 as
3m
3m
3m
3m
2/3
2/3 2/3 ⎛ Ac 2 / p 2 ⎞ ⎛ p1 ⎞ V2 (a / n ) Ac 2 R h22/ 3 ⎛⎜ R h 2 ⎞⎟ ⎟ ⎟⎟ = = = ⎜⎜ = ⎜⎜ ⎟ V1 (a / n ) Ac1 R h21/ 3 ⎜⎝ R h1 ⎟⎠ ⎝ p2 ⎠ ⎝ Ac1 / p1 ⎠ where p is the wetted perimeter. Substituting, 2/3 2/3 2/3 V2 ⎛ p 2 ⎞ ⎛ 6×3m ⎞ ⎛ 3⎞ ⎟⎟ = ⎜⎜ =⎜ =⎜ ⎟ = 1.31 (31% increase) ⎟ ⎝ 4×3m ⎠ ⎝2⎠ V1 ⎝ p 2 ⎠
Discussion This is a very significant increase, and shows the importance of eliminating unnecessary surfaces in flow systems, including pipe flow.
13-57 Solution The flow of water in a trapezoidal channel made of unfinished-concrete is considered. For given flow rate and bottom slope, the flow depth is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel of unfinished concrete is n = 0.014 (Table 13-1).
Analysis
From geometric considerations, the flow area, wetted perimeter, and hydraulic radius are 5 m + 5 m + 2h Ac = h = (5 + h )h 2
p = (5 m) + 2h / sin 45° = 5 + 2.828h Rh =
h 45°
45°
Ac (5 + h )h = p 5 + 2h / sin 45°
5m
Substituting the given quantities into Manning’s equation, (5 + h)h a 1 m 1/3 / s ⎛ ⎞ V = Ac R h2 / 3 S 01 / 2 → 25 m 3 /s = (5 + h)h⎜ ⎟ n 0.014 ⎝ 5 + 2h / sin 45° ⎠
2/3
(tan 1°) 1 / 2
It gives the flow depth to be h = 0.685 m. Discussion Non-linear equations frequently arise in the solution of open channel flow problems. They are best handled by equation solvers such as EES.
13-21
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Chapter 13 Open-Channel Flow 13-58 Solution The flow of water in a weedy excavated trapezoidal channel is considered. For given flow rate and bottom slope, the flow depth is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for the channel is given to be n = 0.030.
Analysis
From geometric considerations, the flow area, wetted perimeter, and hydraulic radius are 5 m + 5 m + 2h Ac = h = (5 + h )h 2 p = (5 m) + 2h / sin 45° = 5 + 2.828h
h 45°
45°
A (5 + h )h Rh = c = p 5 + 2h / sin 45°
5m
Substituting the given quantities into Manning’s equation, (5 + h )h a 1 m 1/3 / s ⎛ ⎞ V = Ac R h2 / 3 S 01 / 2 → 25 m 3 /s = (5 + h )h⎜ ⎟ n 0.030 ⎝ 5 + 2h / sin 45° ⎠
2/3
(tan 1°) 1 / 2
It gives the flow depth to be y = 1.07 m. Discussion Note that as the Manning coefficient increases because of the increased surface roughness of the channel, the flow depth required to maintain the same flow rate also increases.
13-59 Solution The flow of water in a V-shaped cast iron channel is considered. For a given flow depth and bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel of cast iron is n = 0.013 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are 2h × h Ac = = h 2 = (1 m) 2 = 1 m 2 p = 2h / sin θ = 2(1 m)/sin45° = 2 2 m 2
Analysis
Rh =
Ac 1m 2 = = 0.3536 m p 2 2 m
The bottom slope of the channel is S 0 = tan 0.5 D = 0.008727 . Then the flow rate can be determined from Manning’s equation to be
45°
45°
h=1m
a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (1 m 2 )(0.3536 m) 2/3 (0.008727) 1 / 2 = 3.59 m 3 s n 0.013
Discussion
Note that the flow rate in a given channel is a strong function of the bottom slope.
13-22
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Chapter 13 Open-Channel Flow 13-60E Solution The flow of water in a rectangular cast iron channel is considered. For given flow rate and bottom slope, the flow depth is to be determined. Assumptions
1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness coefficient is constant.
Properties
Manning coefficient for a cast iron open channel is n = 0.013 (Table 13-1).
From the geometry, the flow area, wetted perimeter, and hydraulic radius are A 6y Rh = c = Ac = by = (6 ft) y = 6 y p = (6 ft) + 2 y = 6 + 2 y p 6 + 2y
Analysis
y
The channel bottom slope is S0 = 1.5/1000 = 0.0015. Substituting the given quantities into Manning’s equation, ⎛ 6y ⎞ a 1.486 ft 1/3 / s ⎟⎟ V = Ac R h2 / 3 S 01 / 2 → 70 ft 3 /s = (6 y )⎜⎜ n 0.013 ⎝ 6 + 2y ⎠
2/3
(0.0015) 1 / 2
b = 6 ft
Solution of the above equation gives the flow depth to be h = 2.24 ft. Discussion Non-linear equations frequently arise in the solution of open channel flow problems. They are best handled by equation solvers such as EES.
13-61 Solution Water is flowing through a channel with nonuniform surface properties. The flow rate and the effective Manning coefficient are to be determined. 6 10 Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The Manning 1 coefficients do not vary along the channel. 2 Concrete channel 2 Analysis The channel involves two parts n1 = 0.014 with different roughness, and thus it is appropriate Light brush to divide the channel into two subsections. The 1.5 n2 = 0.050 flow rate for each subsection can be determined 2 from the Manning equation, and the total flow rate can be determined by adding them up. The flow area, perimeter, and hydraulic radius for each subsection and the entire channel are: A 18 m 2 Subsection 1: Ac1 = 18 m 2 , R h1 = c1 = p1 = 9 m , = 2.00 m p1 9m
Subsection 2:
Ac 2 = 20 m 2 ,
p 2 = 12 m ,
Rh2 =
Ac 2 20 m 2 = = 1.67 m p2 12 m
Ac 38 m 2 = = 1.81 m p 21 m Applying the Manning equation to each subsection, the total flow rate through the channel is determined to be
Entire channel: Ac = 38 m 2 ,
p = 21 m ,
Rh =
⎛ (18 m 2 ) (2 m) 2/3 (20 m 2 ) (1.67 m) 2/3 ⎞ a a 1/2 3 + V = V1 + V 2 = A1 R12 / 3 S01 / 2 + A1 R12 / 3 S01 / 2 = (1m1/3 /s ) ⎜ ⎟ (0.002) = 116 m /s n1 n1 0.014 0.05 ⎝ ⎠
Knowing the total flow rate, the effective Manning coefficient for the entire channel can be determined from the Manning equation to be n eff =
aAc R h2 / 3 S 01 / 2 (1 m1/3 / s)(38 m 2 )(1.81 m) 2/3 (0.002)1 / 2 = = 0.0217 V 116 m 3 / s
Discussion The effective Manning coefficient neff lies between the two n values as expected. The weighted average of the Manning coefficient of the channel is nave=(n1p1+ n2p2)/p = 0.035, which is quite different than neff. Therefore, using a weighted average Manning coefficient for the entire channel may be tempting, but it would not be accurate.
13-23
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Chapter 13 Open-Channel Flow 13-62 Solution Waters flows in a partially filled circular channel made of finished concrete. For a given flow depth and bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for an open channel made of finished concrete is n = 0.012 (Table 13-1).
Analysis
From geometric considerations,
cos θ =
R − y 0.5 − 0.25 = = 0.5 R 0.5
→
θ = 60° = 60
2π π = 360 3
Ac = R 2 (θ − sin θ cos θ ) = (0.5 m) 2 [π / 3 − sin(π / 3) cos(π / 3)] = 0.1535 m 2
A π / 3 − sin(π / 3) cos(π / 3) θ − sin θ cos θ Rh = c = R= (0.5 m) = 0.1466 m p 2θ 2π / 3
R = 0.5 m
θ y = 0.25 m
Then the flow rate can be determined from Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (0.1535 m 2 )(0.1466 m) 2/3 (0.002)1 / 2 = 0.159 m 3 s n 0.012
Discussion
Note that the flow rate in a given channel is a strong function of the bottom slope.
13-24
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Chapter 13 Open-Channel Flow
13-63
Solution The previous problem is reconsidered. By varying the flow depth-to-radius ratio from 0.1 to 1.9 for a fixed value of flow area, it is the to be shown that the best hydraulic cross section occurs when the circular channel is half-full, and the results are to be plotted.
The EES Equations window is printed below, along with the tabulated and plotted results.
0.17 0.165 0.16 0.155 0.15
3
a=1 n=0.012 s=0.002 Ac=0.1536 “Flow area kept constant” ratio=y/R “This ratio is varied from 0.1 to 1.9” bdeg=arcsin((R-y)/R) tetadeg=90-bdeg teta=tetadeg*2*pi/360 Ac=R^2*(teta-sin(tetadeg)*cos(tetadeg)) p=2*teta*R Rh=Ac/p Vdot=(a/n)*Ac*Rh^(2/3)*SQRT(s)
V, m /s
Analysis
0.145 0.14 0.135 0.13
Depth-to-radius ratio, y/R 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9
Channel radius, R, m 1.617 0.969 0.721 0.586 0.500 0.440 0.396 0.362 0.335 0.313 0.295 0.279 0.267 0.256 0.247 0.239 0.232 0.227 0.223
0.125 0
Flow rate, V ,m3/s 0.1276 0.1417 0.1498 0.1553 0.1592 0.1620 0.1639 0.1652 0.1659 0.1661 0.1659 0.1653 0.1642 0.1627 0.1607 0.1582 0.1550 0.1509 0.1453
0.4
0.8
1.2
1.6
2
Depth-to-radius ratio, y/R
R
θ y
Discussion The depth-to-radius ratio of y/R = 1 corresponds to a half-full circular channel, and it is clear from the table and the chart that, for a fixed flow area, the flow rate becomes maximum when the channel is half-full.
13-25
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Chapter 13 Open-Channel Flow 13-64 Solution Water is to be transported uniformly in a clean-earth trapezoidal channel. For a specified flow rate, the required elevation drop per km channel length is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
Manning coefficient for the clean-earth lined open channel is n = 0.022 (Table 13-1).
Analysis
The flow area, wetted perimeter, and hydraulic radius of the channel are (1.5 + 1.5 + 2) m Ac = (1 m) = 2.5 m 2 2 p = (1.5 m) + 2 (1 m) 2 + (1 m) 2 = 4.328 m
y=1m Slope 1:1
Ac 2.5 m 2 = = 0.5776 m p 4.328 m
Rh =
b = 1.5 m
Substituting the given quantities into Manning’s equation, a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 → 8 m 3 /s = (2.5 m 2 )(0.5776 m) 2/3 S 01 / 2 n 0.022 It gives the slope to be S0 = 0.0103. Therefore, the elevation drop Δz across a pipe length of L = 1 km must be Δz = S 0 L = 0.0103(1000 m) = 10.3 m
Discussion Note that when transporting water through a region of fixed elevation drop, the only way to increase the flow rate is to use a channel with a larger cross-section.
13-65 Solution A water draining system consists of three circular channels, two of which draining into the third one. If all channels are to run half-full, the diameter of the third channel is to be determined. Assumptions 1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the channel. 4 Losses at the junction are negligible.
The Manning coefficient for asphalt lined open channels is n = 0.016 (Table 13-1).
Properties Analysis
The flow area, wetted perimeter, and hydraulic radius of the two pipes upstream are π (0.6 m) 2 2πR 2π (0.6 m) p= = = 0.5655 m 2 = = 1.885 m 2 2 2 πR 2 / 2 R 0.6 m = = = = 0.30 m 2 2 πR
2
πR 2 Ac Rh = P Then the flow rate through the 2 pipes becomes, from Manning’s equation, Ac =
R = 0.6 m
θ = π/2
a 1 m 1/3 / s V = 2 Ac R h2 / 3 S 01 / 2 = 2 (0.5655 m 2 )(0.30 m) 2/3 (0.0015) 1 / 2 = 1.227 m 3 /s n 0.016
The third channel is half-full, and the flow rate through it remains the same. Noting that the flow area is πR2/2 and the hydraulic radius is R/2, we have 1.227 m 3 /s =
1 m 1/3 / s (πR 2 / 2 m 2 )( R / 2 m) 2/3 (0.0015) 1 / 2 0.016
Solving for R gives R = 0.778 m. Therefore, the diameter of the third channel is D3 = 1.56 m . Discussion
Note that if the channel diameter were larger, the channel would have been less than half full.
13-26
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Chapter 13 Open-Channel Flow 13-66 Solution Water is transported in an asphalt lined open channel at a specified rate. The dimensions of the best crosssection for various geometric shapes are to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties The Manning coefficient for asphalt lined open channels is n = 0.016 (Table 13-1). Analysis (a) Circular channel of Diameter D: Best cross-section occurs when the channel is half-full, and thus the flow area is πD2/8 and the hydraulic radius is a D/4. Then from Manning’s equation, V = Ac R h2 / 3 S 01 / 2 , n 1/3 1 m / s 4 m 3 /s = (πD 2 / 8 m 2 )( D / 4 m) 2/3 (0.0015) 1 / 2 0.016
D
which gives D = 2.42 m. (b) Rectangular channel of bottom width b: For best cross-section, y = b/2. Then Ac = yb = b2/2 and Rh = b/4. From the Manning equation, 1 m 1/3 / s 2 4 m 3 /s = (b / 2 m 2 )(b / 4 m) 2/3 (0.0015) 1 / 2 0.016
y = b/2
which gives b = 2.21 m, and y = b/2 = 1.11 m. b
(c) Trapezoidal channel of bottom width b: For best cross-section, θ = 60° and y = b 3 / 2 . Then, Ac = y (b + b cos θ ) = 0.5 3b 2 (1 + cos 60°) = 0.75 3b 2 , y 3 b . From the Manning equation, = 2 4 1 m 1/3 / s 4 m 3 /s = (0.75 3b 2 m 2 )( 3b / 4 m) 2/3 (0.0015) 1 / 2 0.016
p = 3b , R h =
y=
b
3 b 2
60° b
which gives b = 1.35 m, and y = 1.17 m and θ = 60°.
Discussion The perimeters for the circular, rectangular, and trapezoidal channels are 3.80 m, 4.42 m, and 4.05 m, respectively. Therefore, the circular cross-section has the smallest perimeter.
13-27
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-67E Solution Water is to be transported in a rectangular channel at a specified rate. The dimensions for the best crosssection if the channel is made of unfinished and finished concrete are to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties The Manning coefficient is n = 0.012 and n = 0.014 for finished and unfinished concrete, respectively (Table 13-1). Analysis For best cross-section of a rectangular cross-section, y = b/2. Then Ac = yb = b2/2 and Rh = b/4. The flow rate is determined from the Manning equation, a V = Ac R h2 / 3 S 01 / 2 , n
y = b/2
(a) Finished concrete, n = 0.012: 800 ft 3 /s =
1.486 ft 1/3 / s 2 (b / 2 ft 2 )(b / 4 ft) 2/3 (0.0005) 1 / 2 0.012
b
It gives b = 15.4 ft, and y = b/2 = 7.68 ft (b) Unfinished concrete, n = 0.014: 800 ft 3 /s =
1.486 ft 1/3 / s 2 (b / 2 ft 2 )(b / 4 ft) 2/3 (0.0005) 1 / 2 0.014
It gives b = 16.3 ft, and y = b/2 = 8.13 ft Discussion
Note that channels with rough surfaces require a larger cross-section to transport the same amount of water.
13-28
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Chapter 13 Open-Channel Flow
13-68
Solution Uniform flow in an asphalt-lined rectangular channel is considered. By varying the depth-to-width ratio from 0.1 to 2 in increments of 0.1 for a fixed value of flow area, it is the to be shown that the best hydraulic cross section occurs when y/b = 0.5, and the results are to be plotted.
The EES Equations window is printed below, along with the tabulated and plotted results. 4.4 4.3 4.2 4.1 4
3
a=1 n=0.016 "Manning coefficient" s=0.003 "Bottom slope is constant" Ac=2 "Flow area remains constant at 2 m2" Ratio=y/b Ac=b*y p=b+2*y Rh=Ac/p "Hydraulic radius" Vdot=(a/n)*Ac*Rh^(2/3)*SQRT(s) "Volume flow rate"
V, m /s
Analysis
3.9 3.8 3.7
Depth-to-width ratio, y/b 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0
Channel width, b, m 4.47 3.16 2.58 2.24 2.00 1.83 1.69 1.58 1.49 1.41 1.35 1.29 1.24 1.20 1.15 1.12 1.08 1.05 1.03 1.00
3.6
Flow rate, V ,m3/s 3.546 4.031 4.221 4.295 4.313 4.301 4.273 4.235 4.192 4.147 4.101 4.054 4.008 3.963 3.919 3.876 3.834 3.794 3.755 3.717
3.5 0
0.4
0.8
1.2
1.6
2
Depth-to-witdth ratio, y/b
y
b
Discussion It is clear from the table and the chart that the depth-to-width ratio of y/b = 0.5 corresponds to the best cross-section for an open channel of rectangular cross-section.
13-29
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow
Gradually and Rapidly Varied Flows and Hydraulic Jump
13-69C Solution
We are to discuss the difference between uniform and nonuniform (varied) flow.
Analysis Both uniform and varied flows are steady, and thus neither involves any change with time at a specified location. In uniform flow, the flow depth y and the flow velocity V remain constant whereas in nonuniform or varied flow, the flow depth and velocity vary in the streamwise direction of the flow. In uniform flow, the slope of the energy line is equal to the slope of the bottom surface. Therefore, the friction slope equals the bottom slope, Sf = S0. In varied flow, however, these slopes are different. Discussion
Varied flows are further classified into gradually varied flow (GVF) and rapidly varied flow (RVF).
13-70C Solution
We are to discuss the differences between GVF and RVF.
Analysis Gradually varied flow (GVF) is characterized by gradual variations in flow depth and velocity (small slopes and no abrupt changes) and a free surface that always remains smooth (no discontinuities or zigzags). Rapidly varied flow (RVF) involves rapid changes in flow depth and velocity. A change in the bottom slope or cross-section of a channel or an obstruction on the path of flow may cause the uniform flow in a channel to become gradually or rapidly varied flow. Analytical relations for the profile of the free surface can be obtained in GVF, but this is not the case for RVF because of the intense agitation. Discussion
In many situations, the shape of the free surface must be solved numerically, even for GVF.
13-71C Solution
We are to analyze a claim that wall shear is negligible in RVF but important in GVF.
Analysis Yes, we agree with this claim. Rapidly varied flows occur over a short section of the channel with relatively small surface area, and thus frictional losses associated with wall shear are negligible compared with losses due to intense agitation and turbulence. Losses in GVF, on the other hand, are primarily due to frictional effects along the channel, and should be considered. Discussion There is somewhat of an analogy here with internal flows. In long pipe sections with entrance lengths and/or gradually changing pipe diameter, wall shear is important. However, in short sections of piping with rapid change of diameter or a blockage or turn, etc (minor loss), friction along the wall is typically negligible compared to other losses.
13-72C Solution
We are to analyze what happens to flow depth in a horizontal rectangular channel during subcritical flow.
Analysis
The flow depth y must (c) decrease in the flow direction.
Discussion
Since the flow is subcritical, there is no possibility of a hydraulic jump.
13-73C Solution
We are to analyze what happens to flow depth in a sloped rectangular channel during subcritical flow.
Analysis
The flow depth y must (a) increase in the flow direction.
Discussion
Since the flow is subcritical, there is no possibility of a hydraulic jump.
13-30
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Chapter 13 Open-Channel Flow 13-74C Solution
We are to analyze what happens to flow depth in a horizontal rectangular channel during supercritical flow.
Analysis
The flow depth y (a) increases in the flow direction.
Discussion
Since the flow is supercritical, this increase in flow depth may occur via a hydraulic jump.
13-75C Solution
We are to analyze what happens to flow depth in a sloped rectangular channel during subcritical flow.
Analysis
The flow depth y (c) decreases in the flow direction.
Discussion
Since the flow is subcritical, there is no possibility of a hydraulic jump.
13-76C Solution flow.
We are to analyze what happens to flow depth in an upward-sloped rectangular channel during supercritical
Analysis
The flow depth y (a) increases in the flow direction.
Discussion
Since the flow is supercritical, this increase in flow depth may occur via a hydraulic jump.
13-77C Solution
We are to determine if it is possible for subcritical flow to undergo a hydraulic jump.
Analysis No. It is impossible for subcritical flow to undergo a hydraulic jump. Such a process would require the head loss hL to become negative, which is impossible. It would correspond to negative entropy generation, which would be a violation of the second law of thermodynamics. Therefore, the upstream flow must be supercritical (Fr1 > 1) for a hydraulic jump to occur. Discussion This is analogous to normal shock waves in gases – the only way a shock wave can occur is if the flow upstream of the shock wave is supersonic with Ma1 > 1 (analogous to supercritical in open-channel flow with Fr1 > 1).
13-78C Solution We are to define the energy dissipation ratio for a hydraulic jump and discuss why a hydraulic jump is sometimes used to dissipate energy. Analysis Hydraulic jumps are often designed in conjunction with stilling basins and spillways of dams in order to waste as much of the mechanical energy as possible to minimize the mechanical energy of the fluid and thus its potential to cause damage. In such cases, a measure of performance of a hydraulic jump is the energy dissipation ratio, which is the fraction of energy dissipated through a hydraulic jump, defined as Dissipation ratio =
Discussion
hL hL hL = = . 2 Es1 y1 + V1 / ( 2 g ) y1 (1 + Fr12 / 2 )
Since the head loss is always positive, the dissipation ratio is also always positive.
13-31
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Chapter 13 Open-Channel Flow 13-79
Solution Water is flowing in an open channel uniformly. It is to be determined whether the channel slope is mild, critical, or steep for this flow. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant.
The Manning coefficient for an open channel with finished concrete surfaces is n = 0.012 (Table 13-1).
Properties
Analysis The cross-sectional area, perimeter, and hydraulic radius are Ac = yb = (1.2 m)(3 m) = 3.6 m 2 p = b + 2 y = 3 m + 2(1.2 m) = 5.4 m
Rh =
Ac 3.6 m 2 = = 0.6667 m p 5.4 m
The flow rate is determined from the Manning equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (3.6 m 2 )(0.6667 m) 2/3 (0.002) 1 / 2 = 10.2 m 3 /s n 0.012 Noting that the flow is uniform, the specified flow rate is the normal depth and thus y = yn = 1.2 m. The critical depth for this flow is ⎛ V 2 yc = ⎜ ⎜ g b2 ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (10.2 m 3 /s) 2 =⎜ ⎜ (9.81 m/s 2 ) (3 m) 2 ⎝
⎞ ⎟ ⎟ ⎠
y = 1.2 m
b=3m
1/3
= 1.06 m
This channel at these flow conditions is classified as mild since y > yc, and the flow is subcritical. Discussion If the flow depth were smaller than 1.06 m, the channel slope would be said to be steep. Therefore, the bottom slope alone is not sufficient to classify a downhill channel as being mild, critical, or steep.
13-80 Solution Water is flowing in an wide brick open channel uniformly. The range of flow depth for which the channel can be classified as “steep” is to be determined. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties
The Manning coefficient for a brick open channel is n = 0.015 (Table 13-1).
y
Analysis The slope of the channel is S 0 = tan α = tan 0.4° = 0.006981 . The hydraulic radius for a wide channel is equal to the flow depth, Rh = y. Now assume the flow in the channel to be critical, The channel flow in this case would be critical slope Sc, and the flow depth would be the critical flow depth, which is determined from Sc =
g n2
→
a 2 y 1c 3
⎛ g n2 yc = ⎜ 2 ⎜a S c ⎝
⎞ ⎟ ⎟ ⎠
3
Substituting, ⎛ g n2 yc = ⎜ 2 ⎜a S c ⎝
3
3
2 2 ⎞ ⎞ ⎛ ⎟ = ⎜ (9.81 m/s )(0.015) ⎟ = 0.0316 m ⎜ (1 m 1 3 / s ) 2 (0.006981) ⎟ ⎟ ⎝ ⎠ ⎠
Therefore, this channel can be classified as steep for uniform flow depths less than yc, i.e., y < 0.0316 m. Discussion Note that two channels of the same slope can be classified as differently (one mild and the other steep) if they have different roughness and thus different values of n.
13-32
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Chapter 13 Open-Channel Flow 13-81E Solution Water is flowing in a rectangular open channel with a specified bottom slope at a specified flow rate. It is to be determined whether the slope of this channel should be classified as mild, critical, or steep. The surface profile is also to be classified for a specified flow depth of 2 m. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant.
The Manning coefficient of a channel with unfinished concrete surfaces is n = 0.014 (Table 13-1).
Properties
Analysis The cross-sectional area, perimeter, and hydraulic radius are Ac = yb = y (12 ft) = 12 y ft 2 p = b + 2 y = 12 ft + 2 y = 12 + 2 y ft
Rh =
Ac 12 y ft 2 = = 0.6667 m p 12 + 2y ft
y
Substituting the known quantities into the Manning equation, ⎛ 12 y ⎞ 1.486 ft 1/3 / s a ⎟⎟ (12 y )⎜⎜ V = Ac R h2 / 3 S 01 / 2 → 300 ft 3 /s = 0.014 n ⎝ 12 + 2 y ⎠
2/3
(tan 0.5°) 1 / 2
b = 12 ft
Solving for the flow depth y gives y = 1.95 ft. The critical depth for this flow is yc =
V 2
gAc2
=
(300 ft 3 / s) 2 (32.2 ft/s 2 )(12 ft × 1.95 ft) 2
= 5.10 ft
This channel at these flow conditions is classified as steep since y < yc, and the flow is supercritical. Alternately, we could solve for Froude number and show that Fr > 1 and reach the same conclusion. The given flow is uniform, and thus y = yn = 1.95 ft. Therefore, the given value of y = 3 ft during development is between yc and yn, and the flow profile is S2 (Table 13-3). Discussion If the flow depth were larger than 5.19 ft, the channel slope would be said to be mild. Therefore, the bottom slope alone is not sufficient to classify a downhill channel as being mild, critical, or steep.
13-82 Solution Water is flowing in a V-shaped open channel with a specified bottom slope at a specified rate. It is to be determined whether the slope of this channel should be classified as mild, critical, or steep. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness of the wetted surface of the channel and thus the friction coefficient are constant. Properties
The Manning coefficient for a cast iron channel is n = 0.013 (Table 13-1).
From geometric considerations, the cross-sectional area, perimeter, and hydraulic radius are A y2 y p = 2 y2 + y2 = 2 2 y Rh = c = Ac = y (2 y ) / 2 = y 2 = p 2 2y 2 2
Analysis
Substituting the known quantities into the Manning equation, a 1 m 1/3 / s 2 ⎛ y ⎞ ⎟⎟ V = Ac R h2 / 3 S 01 / 2 → 3 m 3 /s = ( y )⎜⎜ n 0.013 ⎝2 2 ⎠
2/3
(0.002) 1 / 2
Solving for the flow depth y gives y = 1.23 m. The critical depth for this flow is (3 m 3 / s) 2 V 2 yc = = = 0.61 m gAc2 (9.81 m/s 2 )(1.23 m) 2
45°
45°
y
This channel at these flow conditions is classified as mild since y > yc, and the flow is subcritical. Discussion If the flow depth were smaller than 0.61 m, the channel slope would be said to be steep. Therefore, the bottom slope alone is not sufficient to classify a downhill channel as being mild, critical, or steep.
13-33
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Chapter 13 Open-Channel Flow 13-83 [Also solved using EES on enclosed DVD] Solution Water at a specified depth and velocity undergoes a hydraulic jump. The depth and Froude number after the jump, the head loss and dissipation ratio, and dissipated mechanical power are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. Properties
The density of water is 1000 kg/m3.
Analysis
(a) The Froude number before the hydraulic jump is
Fr1 =
V1 gy1
=
9 m/s (9.81 m/s 2 )(1.2 m)
hL
= 2.62
which is greater than 1. Therefore, the flow is supercritical before the jump. The flow depth, velocity, and Froude number after the jump are
Energy line
V1 = 9 m/s V2 FIGURE 13-51 y2 Control volume that encloses the y1 = 1.2 m hydraulic jump. (1)
(2)
y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ = 0.5(1.2 m)⎛⎜ − 1 + 1 + 8 × 2.62 2 ⎞⎟ = 3.89 m ⎝ ⎠ ⎝ ⎠ V2 = Fr2 =
y1 1.2 m V1 = (9 m/s ) = 2.78 m/s y2 3.89 m V2 gy 2
=
2.78 m/s (9.81 m/s 2 )(3.89 m)
= 0.449
(b) The head loss is determined from the energy equation to be h L = y1 − y 2 +
V12 − V 22 (9 m/s) 2 - (2.78 m/s) 2 = (1.2 m) - (3.89 m) + = 1.05 m 2g 2(9.81 m/s 2 )
The specific energy of water before the jump and the dissipation ratio are E s1 = y1 +
V12 (9 m/s) 2 = (1.2 m) + = 5.33 m 2g 2(9.81 m/s 2 )
Dissipation ratio =
h L 1.04 m = = 0.195 E s1 5.33 m
Therefore, 19.5% of the available head (or mechanical energy) of the liquid is wasted (converted to thermal energy) as a result of frictional effects during this hydraulic jump. (c) The mass flow rate of water is m = ρV = ρby1V1 = (1000 kg/m 3 )(1.2 m)(8 m)(9 m/s) = 86,400 kg/s
Then the dissipated mechanical power becomes ⎛ 1N E dissipated = m gh L = (86,400 kg/s)(9.81 m/s 2 )(1.04 m)⎜⎜ 1 kg ⋅ m/s 2 ⎝
⎞ ⎟ = 881,000 Nm/s = 881 kW ⎟ ⎠
Discussion The results show that the hydraulic jump is a highly dissipative process, wasting 881 kW of power production potential in this case. That is, if the water is routed to a hydraulic turbine instead of being released from the sluice gate, up to 881 kW of power could be produced.
13-34
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Chapter 13 Open-Channel Flow 13-84 Solution Water at a specified depth and velocity undergoes a hydraulic jump. The head loss associated with this process is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal.
The Froude number before the hydraulic jump is
Analysis
V1
Fr1 =
gy1
=
12 m/s (9.81 m/s 2 )(0.35 m)
= 6.476 , which is
greater than 1. Therefore, the flow is indeed supercritical before the jump. The flow depth, velocity, and Froude number after the jump are y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ = 0.5(0.35 m)⎛⎜ − 1 + 1 + 8 × 6.476 2 ⎞⎟ = 3.035 m ⎝ ⎠ ⎝ ⎠ V2 =
hL
y1 0.35 m V1 = (12 m/s ) = 1.384 m/s y2 3.035 m
Fr2 =
V2 gy 2
1.384 m/s
=
2
V1 = 12 m/s V2 FIGURE 13-51 y2 Control volume that encloses the y1 = 0.35 m hydraulic jump.
= 0.2536
(9.81 m/s )(3.035 m)
The head loss is determined from the energy equation to be hL = y1 − y2 +
Discussion process.
V12
− V22
2g
= (0.35 m) - (3.035 m) +
Energy line
(1) 2
(2)
2
(12 m/s) - (1.384 m/s) = 4.56 m 2(9.81 m/s 2 )
The results show that the hydraulic jump is a highly dissipative process, wasting 4.56 m of head in the
13-35
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-85 Solution The increase in flow depth during a hydraulic jump is given. The velocities and Froude numbers before and after the jump, and the energy dissipation ratio are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. hL Energy Analysis The Froude number before the jump is determined from line
y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ → 3 m = 0.5 × (0.6 m)⎛⎜ − 1 + 1 + 8Fr 2 ⎞⎟ ⎝ ⎠ ⎝ ⎠
which gives Fr1 = 3.873. Then, V1 = Fr1 gy1 = 3.873 (9.81 m/s 2 )(0.6 m) = 9.40 m/s V2 =
Fr2 =
V
V2 FIGURE 13-51 y =3m Control volume that2encloses the y1 = 0.6 m hydraulic jump. 1
(1)
(2)
y1 0.6 m V1 = (9.40 m/s ) = 1.88 m/s y2 3m V2 gy 2
=
1.88 m/s (9.81 m/s 2 )(3 m)
= 0.347
The head loss is determined from the energy equation to be h L = y1 − y 2 +
V12 − V 22 (9.40 m/s) 2 - (1.88 m/s) 2 = (0.6 m) - (3 m) + = 1.92 m 2g 2(9.81 m/s 2 )
The specific energy of water before the jump and the dissipation ratio are E s1 = y1 +
V12 (9.40 m/s) 2 = (0.6 m) + = 5.10 m 2g 2(9.81 m/s 2 )
Dissipation ratio =
hL 1.92 m = = 0.376 E s1 5.10 m
Therefore, 37.6% of the available head (or mechanical energy) of water is wasted (converted to thermal energy) as a result of frictional effects during this hydraulic jump. The results show that the hydraulic jump is a highly dissipative process, wasting over one-third of the Discussion available head.
13-36
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-86 Solution Water flowing in a wide channel at a specified depth and flow rate undergoes a hydraulic jump. The mechanical power wasted during this process is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. Properties Analysis
The density of water is 1000 kg/m3.
hL
Average velocities before and after the jump are 70 m 3 /s V1 = = 14 m/s (10 m)( 0.5 m)
V
V2 FIGURE 13-51 y =4m Control volume that2encloses the y1 = 0.5 m hydraulic jump.
3
70 m /s = 1.75 m/s (10 m)(4 m) The head loss is determined from the energy equation to be V2 =
h L = y1 − y 2 +
Energy line
1
(1)
(2)
V12 − V 22 (14 m/s) 2 - (1.75 m/s) 2 = (0.5 m) - (4 m) + = 6.33 m 2g 2(9.81 m/s 2 )
The mass flow rate of water is m = ρV = (1000 kg/m 3 )(70 m 3 /s) = 70,000 kg/s
Then the dissipated mechanical power becomes ⎛ 1 kN E dissipated = m gh L = (70,000 kg/s)(9.81 m/s 2 )(6.33 m)⎜⎜ 2 ⎝ 1000 kg ⋅ m/s
⎞ ⎟ = 4350 kNm/s = 4.35 MW ⎟ ⎠
Discussion The results show that the hydraulic jump is a highly dissipative process, wasting 4.35 MW of power production potential in this case.
13-37
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Chapter 13 Open-Channel Flow 13-87 Solution The flow depth and average velocity of water after a hydraulic jump are measured. The flow depth and velocity before the jump as well as the fraction of mechanical energy dissipated are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. hL Energy Analysis The Froude number after the hydraulic jump is V2
Fr2 =
gy 2
=
3 m/s (9.81 m/s 2 )(2 m)
line
= 0.6773
It can be shown that the subscripts in the relation y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ are interchangeable. Thus, ⎝ ⎠ ⎛ y1 = 0.5 y 2 ⎜ − 1 + 1 + 8Fr22 ⎞⎟ = 0.5(2 m)⎛⎜ − 1 + 1 + 8 × 0.6773 2 ⎞⎟ = 1.16 m ⎠ ⎝ ⎠ ⎝ y2 2m V1 = V2 = (3 m/s ) = 5.17 m/s y1 1.161 m The Froude number before the jump is Fr1 =
V1 gy1
=
5.17 m/s (9.81 m/s 2 )(1.16 m)
V 2= 3 m/s FIGURE 13-51 y =2m Control volume that2encloses the y1 hydraulic jump.
V
1
(1)
(2)
= 1.53
which is greater than 1. Therefore, the flow is indeed supercritical before the jump. The head loss is determined from the energy equation to be h L = y1 − y 2 +
V12 − V 22 (5.17 m/s) 2 - (3 m/s) 2 = (1.16 m) - (2 m) + = 0.0636 m 2g 2(9.81 m/s 2 )
The specific energy of water before the jump and the dissipation ratio is E s1 = y1 +
V12 (5.17 m/s) 2 = (1.16 m) + = 2.52 m 2g 2(9.81 m/s 2 )
Dissipation ratio =
Discussion
hL 0.0636 m = = 0.025 2.52 m E s1
Note that this is a “mild” hydraulic jump, and only 2.5% of the available energy is wasted.
13-38
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-88E Solution Water at a specified depth and velocity undergoes a hydraulic jump, and dissipates a known fraction of its energy. The flow depth, velocity, and Froude number after the jump and the head loss associated with the jump are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. hL Energy Analysis The Froude number before the hydraulic jump is line V 40 ft/s Fr1 = 1 = = 4.984 gy1 (32.2 m/s 2 )(2 ft) V1 = 40 ft/s V2 FIGURE 13-51 y2 which is greater than 1. Therefore, the flow is indeed Control volume that encloses the y1 = 2 ft supercritical before the jump. The flow depth, velocity, and hydraulic jump. Froude number after the jump are (1) (2) y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ = 0.5(2 ft)⎛⎜ − 1 + 1 + 8 × 4.984 2 ⎞⎟ = 13.1 ft ⎠ ⎝ ⎠ ⎝ y 2 ft V 2 = 1 V1 = (40 ft/s ) = 6.09 ft/s y2 13.1 ft Fr2 =
V2 gy 2
=
6.091 ft/s (32.2 ft/s 2 )(13.13 m)
= 0.296
The head loss is determined from the energy equation to be V 2 − V 22 (40 ft/s) 2 - (6.09 ft/s) 2 = (2 ft) - (13.1 ft) + = 13.2 ft h L = y1 − y 2 + 1 2g 2(32.2 ft/s 2 ) Discussion process.
The results show that the hydraulic jump is a highly dissipative process, wasting 13.2 ft of head in the
Flow Control and Measurement in Channels
13-89C Solution We are to draw a flow depth-specific energy diagram for several types of flow.
Subcritical flow
y 1
Analysis On the figure, diagram 1-2a is for frictionless gate, 1-2b is for sluice gate with free outflow, and 1-2b-2c is for sluice gate with drown outflow, including the hydraulic jump back to subcritical flow.
• 2c (c) Drown outflow
Discussion A plot of flow depth as a function of specific energy, as shown here, is quite useful in the analysis of varied open-channel flow because the states upstream and downstream of a change must jump between the two branches.
•
(a) Frictionless gate
• • 2b
Supercritical flow
2a
Es
Es1 = Es2a
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PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-90C Solution
We are to define the discharge coefficient for sluice gates, and discuss some typical values.
Analysis For sluice gates, the discharge coefficient Cd is defined as the ratio of the actual velocity through the gate to the maximum velocity as determined by the Bernoulli equation for the idealized frictionless flow case. For ideal flow, Cd = 1. Typical values of Cd for sluice gates with free outflow are in the range of 0.55 to 0.60. Discussion
Actual values of the discharge coefficient must be less than one or else the second law would be violated.
13-91C Solution
We are to discuss how flow rate is measured with a broad-crested weir.
Analysis The operation of broad crested weir is based on blocking the flow in the channel with a rectangular block, and establishing critical flow over the block. Then the flow rate is determined by measuring flow depths. Discussion
This technique is quite obtrusive, but requires no special measuring equipment or probes.
13-92C Solution
We are to analyze whether the free surface of flow over a bump will increase, decrease, or remain constant.
Analysis
In the case of subcritical flow, the flow depth y will decrease during flow over the bump.
Discussion This may be contrary to our intuition at first, but if we think in terms of increasing velocity and decreasing pressure over the bump (a Bernoulli type of analysis), it makes sense that the surface will decrease over the bump.
13-93C Solution
We are to analyze what happens in subcritical flow over a bump when the bump height increases.
Analysis When the specific energy reaches its minimum value, the flow is critical, and the flow at this point is said to be choked. If the bump height is increased even further, the flow remains critical and thus choked. The flow will not become supercritical. Discussion This is somewhat analogous to compressible flow in a converging nozzle – the flow cannot become supersonic at the nozzle exit unless there is a diverging section of the nozzle downstream of the throat.
13-94C Solution
We are to define and classify sharp-crested weirs.
Analysis A sharp-crested weir is a vertical plate placed in a channel that forces the fluid to flow through an opening to measure the flow rate. They are characterized by the shape of the opening. For example, a weir with a triangular opening is referred to as a triangular weir. Discussion Similar to the broad-crested weir, this type of flow measurement is quite obtrusive, but requires no special measuring equipment or probes.
13-40
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Chapter 13 Open-Channel Flow 13-95
Solution Water is released from a reservoir through a sluice gate into an open channel. For specified flow depths, the rate of discharge is to be determined. Sluice gate Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible.
The depth ratio y1/a and the contraction coefficient y2/a are
Analysis
y1 14 m = = 14 a 1m
and
y2 3 m = =3 a 1m
y1 = 14 m
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.59. Then the discharge rate becomes
y2 = 3 m
a=1m V = C d ba 2 gy1 = 0.59 (5 m)(1 m) 2 (9.81 m/s 2 )(14 m) = 48.9 m 3 /s
Discussion Discharge coefficient is the same as free flow because of small depth ratio after the gate. So, the flow rate would not change if it were not drowned.
13-96 Solution Water flowing in a horizontal open channel encounters a bump. It will be determined if the flow over the bump is choked.
Depression over the bump
Assumptions 1 The flow is steady. 2 Frictional effects are negligible so that there is no dissipation of mechanical energy. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis
Fr1 =
y1=1.2 =0.80mm =1.2 m/s m/s VV11 =2.5
The upstream Froude number and the critical depth are 2.5 m/s = = 0.729 gy1 (9.81 m 2 /s)(1.2 m)
Δzb==0.22 0.15 m
V1
1/ 3
1/ 3
1/ 3
y2
Bump 1/ 3
⎛ (by V ) 2 ⎞ ⎛ y 2V 2 ⎞ ⎛ (1.2 m) 2 (2.5 m / s) 2 ⎞ ⎛ V 2 ⎞ ⎟ = 0.972 m y c = ⎜⎜ 2 ⎟⎟ = ⎜⎜ 1 21 ⎟⎟ = ⎜⎜ 1 1 ⎟⎟ = ⎜⎜ ⎟ 9.81 m/s 2 ⎠ ⎝ ⎝ gb ⎠ ⎝ gb ⎠ ⎝ g ⎠ The flow is subcritical since Fr < 1, and the flow depth decreases over the bump. The upstream, over the bump, and critical specific energies are V2 (2.5 m/s) 2 E s1 = y1 + 1 = (1.2 m) + = 1.52 m 2g 2(9.81 m/s 2 ) E s 2 = E s1 − Δz b = 1.52 − 0.22 = 1.30 m 3 y c = 1.46 m 2 We have an interesting situation: The calculations show that Es2 < Ec. That is, the specific energy of the fluid decreases below the level of energy at the critical point, which is the minimum energy, and this is impossible. Therefore, the flow at specified conditions cannot exist. The flow is choked when the specific energy drops to the minimum value of 1.46 m, which occurs at a bump-height of Δz b,max = E s1 − E c = 1.52 − 1.46 = 0.06 m .
Ec =
Discussion A bump-height over 6 cm results in a reduction in the flow rate of water, or a rise of upstream water level. Therefore, a 22-cm high bump alters the upstream flow. On the other hand, a bump less than 6 cm high will not affect the upstream flow.
13-41
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Chapter 13 Open-Channel Flow 13-97 Solution Water flowing in a horizontal open channel encounters a bump. The change in the surface level over the bump and the type of flow (sub- or supercritical) over the bump are to be determined.
Rise over the bump
Assumptions 1 The flow is steady. 2 Frictional effects are negligible so that there is no dissipation of mechanical energy. 3 The channel is sufficiently wide so that the end effects are negligible.
y1= =0.80 0.8 m m = 8 m/s m/s VV11 =1.2
The upstream Froude number and the critical depth are V 8 m/s Fr1 = 1 = = 2.856 gy1 (9.81 m 2 /s)(0.8 m)
Analysis
⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (by V ) 2 = ⎜⎜ 1 21 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ y 2V 2 = ⎜⎜ 1 1 ⎝ g
⎞ ⎟ ⎟ ⎠
1/ 3
y2 Δzb==0.30 0.15 m
Bump
⎛ (0.8 m) 2 (8 m / s) 2 = ⎜⎜ 9.81 m/s 2 ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
= 1.61 m
The upstream flow is supercritical since Fr > 1, and the flow depth increases over the bump. The upstream, over the bump, and critical specific energies are E s1 = y1 +
V12 (8 m/s) 2 = (0.8 m) + = 4.06 m 2g 2(9.81 m/s 2 )
E s 2 = E s1 − Δz b = 4.06 − 0.30 = 3.76 m 3 y c = 2.42 m 2 The flow depth over the bump is determined from
Ec =
y 23 − ( E s1 − Δz b ) y 22 +
V12 2 y1 = 0 → 2g
y 23 − (4.06 − 0.30 m ) y 22 +
(8 m/s) 2 2
2(9.81 m/s )
(0.80 m) 2 = 0
Using an equation solver, the physically meaningful root of this equation is determined to be 0.846 m. Therefore, there is a rise of Rise over bump = y 2 − y1 + Δz b = 0.846 − 0.80 + 0.30 = 0.346 m
over the surface relative to the upstream water surface. The specific energy decreases over the bump from, 4.06 to 3.76 m, but it is still over the minimum value of 2.42 m. Therefore, the flow over the bump is still supercritical. Discussion The actual value of surface rise may be different than 4.6 cm because of frictional effects that are neglected in this simplified analysis.
13-42
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Chapter 13 Open-Channel Flow 13-98 Solution The flow rate in an open channel is to be measured using a sharp-crested rectangular weir. For a measured value of flow depth upstream, the flow rate is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible.
V1
y1 = 2.2 m
Analysis The weir head is H = y1 − Pw = 2.2 − 0.75 = 1.45 m
Pw = 0.75 m
The discharge coefficient of the weir is H 1.45 m C wd ,rec = 0.598 + 0.0897 = 0.598 + 0.0897 = 0.771 Pw 0.75 m
Sharp-crested rectangular weir
The condition H/Pw < 2 is satisfied since 1.45/0.75 = 1.93. Then the water flow rate through the channel becomes 2 2 3/ 2 Vrec = Cwd ,rec b 2 g H 3 / 2 = ( 0.7714 ) ( 4 m ) 2 ( 9.81 m/s 2 ) (1.45 m ) = 15.9 m3 /s 3 3
Discussion
The upstream velocity and the upstream velocity head are V1 =
V 15.9 m 3 /s = = 1.81 m/s and by1 (4 m)(2.2 m)
V12 (1.81 m/s) 2 = = 0.167 m respectively. This is 11.5% of the weir head, which is significant. When the upstream 2 g 2(9.81 m/s 2 ) velocity head is considered, the flow rate becomes 18.1 m3/s, which is about 14 percent higher than the value determined above. Therefore, it is good practice to consider the upstream velocity head unless the weir height Pw is very large relative to the weir head H.
13-99 Solution The flow rate in an open channel is to be measured using a sharp-crested rectangular weir. For a measured value of flow depth upstream, the flow rate is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis The weir head is H = y1 − Pw = 2.2 − 1.0 = 1.2 m . The discharge coefficient of the weir is H 1.2 m = 0.598 + 0.0897 = 0.7056 C wd ,rec = 0.598 + 0.0897 Pw 1.0 m
The condition H/Pw < 2 is satisfied since 1.2/1.0 = 1.20. Then the water flow rate through the channel becomes
V1
y1 = 2.2 m Pw = 1 m
Sharp-crested rectangular weir
2 2 3/ 2 Vrec = Cwd ,rec b 2 g H 3 / 2 = ( 0.7056 ) ( 4 m ) 2 ( 9.81 m/s 2 ) (1.2 m ) = 11.0 m3 /s 3 3 V 11.0 m 3 /s Discussion The upstream velocity and the upstream velocity head are V1 = = = 1.25 m/s and by1 (4 m)(2.2 m) V12 (1.25 m/s) 2 = = 0.079 m , respectively. This is 6.6% of the weir head, which may be significant. When the upstream 2 g 2(9.81 m/s 2 ) velocity head is considered, the flow rate becomes 11.9 m3/s, which is about 8 percent higher than the value determined above. Therefore, it is good practice to consider the upstream velocity head unless the weir height Pw is very large relative to the weir head H.
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Chapter 13 Open-Channel Flow 13-100 Solution Water flowing over a sharp-crested rectangular weir is discharged into a channel where uniform flow conditions are established. The maximum slope of the downstream channel to avoid hydraulic jump is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Properties Manning coefficient for an open channel of unfinished concrete is n = 0.014 (Table 13-1).
V1
y1 = 3 m
Analysis The weir head is H = y1 − Pw = 3.0 m − 2.0 m = 1.0 m . The condition H/Pw < 2 is satisfied since 1.0/2.0 = 0.5. The discharge coefficient of the weir is H 1.0 m C wd ,rec = 0.598 + 0.0897 = 0.598 + 0.0897 = 0.6429 Pw 2.0 m
Pw = 2 m
Sharp-crested rectangular weir
Then the water flow rate through the channel per meter width (i.e., taking b = 1 m) becomes 2 2 3/ 2 Vrec = Cwd ,rec b 2 g H 3 / 2 = ( 0.6429 ) (1 m ) 2 ( 9.81 m/s 2 ) (1.0 m ) = 1.898 m 3 /s 3 3
To avoid hydraulic jump, we must avoid supercritical flow in the channel. Therefore, the bottom slope should not be higher than the critical slope, in which case the flow depth becomes the critical depth, ⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (1.898 m 3 /s) 2 ⎞ ⎟ = ⎜⎜ 2 2 ⎟ ⎝ (9.81 m/s )(1 m ) ⎠
1/ 3
= 0.7162 m
Noting that the hydraulic radius of a wide channel is equal to the flow depth, the bottom slope is determined from the Manning equation to be a 1 m 1/3 / s (0.7162 × 1 m 2 )(0.7162 m) 2/3 S 01 / 2 V = Ac R h2 / 3 S 01 / 2 → 1.898 m 3 /s = 0.014 n
Solution gives the slope to be S0 = 0.00215. Therefore, S0, max = 0.00215. Discussion For a bottom slope smaller than calculated value, downstream channel would have a mild slope, that will force the flow to remain subcritical.
13-101E Solution The flow rate in an open channel is to be measured using a sharp-crested rectangular weir. For specified upper limits of flow rate and flow depth, the appropriate height of the weir is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis C wd ,rec
The weir head is H = y1 − Pw = 5 − Pw . The discharge coefficient of the weir is 5 − Pw H = 0.598 + 0.0897 = 0.598 + 0.0897 Pw Pw
The water flow rate through the channel can be expressed as 2 Vrec = C wd ,rec b 2 g H 3 / 2 3 Substituting the known quantities, ⎛ 5 − Pw 150 ft 3 /s = ⎜⎜ 0.598 + 0.0897 Pw ⎝
⎞2 ⎟ (10 ft) 2(32.2 ft/s 2 ) (5 − Pw ) 3 / 2 ⎟3 ⎠
V1
y1 = 5 ft Pw
Sharp-crested rectangular weir
Solution of the above equation yields the weir height as Pw = 2.46 ft. Discussion
Nonlinear equations of this kind can be solved easily using equation solvers like EES.
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Chapter 13 Open-Channel Flow 13-102 Solution The flow of water in a wide channel with a bump is considered. The flow rate of water without the bump and the effect of the bump on the flow rate for the case of a flat surface are to be determined. Assumptions 1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the channel. 4 The channel is sufficiently wide so that the end effects are negligible. 5 Frictional effects during flow over the bump are negligible.
y1 = 2 m
Properties Manning coefficient for an open channel of unfinished concrete is n = 0.014 (Table 13-1). Analysis For a wide channel, the hydraulic radius is equal to the flow depth, and thus Rh = 2 m. Then the flow rate before the bump per m width (i.e., b = 1 m) can be determined from Manning’s equation to be
y2
Δzb = 15 cm Bump Slope = 0.0022
a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (1× 2 m 2 )(2 m) 2/3 (0.0022) 1 / 2 = 10.64 m 3 /s n 0.014
The average flow velocity is V =
V 10.64 m 3 /s = = 5.32 m/s . Ac 1× 2 m 2
When a bump is placed, it is said that the flow depth remains the same and there is no rise/drop, and thus y 2 = y1 − Δz b . But the energy equation is given as E s 2 = E s1 − Δz b
→
y2 +
V 22 V2 V 22 V12 = y1 + 1 − Δz b → = 2g 2g 2g 2g
since y 2 = y1 − Δz b , and thus V1 = V2. But from the continuity equation y 2V 2 = y1V1 , this is possible only if the flow depth over the bump remains constant, i.e., y1 = y2, which is a contradiction since y2 cannot be equal to both y1 and y1 − Δz b while Δz b remains nonzero. Therefore, the second part of the problem can have no solution since it is physically impossible. Discussion Note that sometimes it is better to investigate whether there is really a solution before spending a lot of time trying to find a solution.
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Chapter 13 Open-Channel Flow 13-103 Solution Uniform subcritical water flow of water in a wide channel with a bump is considered. For critical flow over the bump, the flow rate of water and the flow depth over the bump are to be determined. Assumptions 1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the channel. 4 The channel is sufficiently wide so that the end effects are negligible. 5 Frictional effects during flow over the bump are negligible.
Critical flow
Properties Manning coefficient for an open channel of unfinished concrete is n = 0.014 (Table 13-1). Analysis Let subscript 1 denote the upstream conditions (uniform flow) in the channel, and 2 denote the critical conditions over the bump. For a wide channel, the hydraulic radius is equal to the flow depth, and thus Rh = y1. Then the flow rate per m width (i.e., b = 1 m) can be determined from Manning’s equation,
y1
y2
Δzb = 15 cm Bump Slope = 0.0022
a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = y1 ( y1 ) 2/3 (0.0022) 1 / 2 = 3.350 y15 / 3 m 3 /s n 0.014
The critical depth corresponding to this flow rate is (note that b = 1 m), ⎛ V 2 ⎞ y 2 = y c = ⎜⎜ 2 ⎟⎟ ⎝ gb ⎠
1/ 3
⎛ (3.350 y15 / 3 ) 2 = ⎜⎜ g ⎝
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ 11.224 y110 / 3 = ⎜⎜ 2 ⎝ 9.81 m/s
⎞ ⎟ ⎟ ⎠
1/ 3
= 1.046 y110 / 9
The average flow velocity is V1 = V / Ac = 3.350 y15 / 3 / y1 = 3.350 y12 / 3 m/s . Also, V12 (3.350 y12 / 3 ) 2 = y1 + = y1 + 0.5720 y14 / 3 2 2g 2(9.81 m/s ) 3 3 E s 2 = E c = y c = (1.046 y110 / 9 ) = 1.569 y110 / 9 2 2 Substituting these two relations into E s 2 = E s1 − Δz b where Δz b = 0.15 m gives E s1 = y1 +
1.569 y110 / 9 = y1 + 0.5720 y14 / 3 − 0.15
Using an equation solver such as EES or an iterative approach, the flow depth upstream is determined to be y1 = 2.947 m Then the flow rate and the flow depth over the bump becomes V = 3.350 y15 / 3 = 3.350(2.947) 5/3 = 20.3 m 3 /s
y 2 = y c = 1.046 y110 / 9 = 1.046(2.947)10 / 9 = 3.48 m
Discussion Note that when critical flow is established and the flow is “choked”, the flow rate calculations become very easy, and it required minimal measurements. Also, V1 = 3.350(2.947) 2 / 3 = 6.89 m/s and Fr1 = V1 / gy1 = (6.89 m/s) / (9.81 m 2 /s)(2.947 m) = 1.28 , and thus the upstream flow is supercritical.
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Chapter 13 Open-Channel Flow 13-104 Solution A sluice gate is used to control the flow rate of water in a channel. For specified flow depths upstream and downstream from the gate, the flow rate of water and the downstream Froude number are to be determined. Assumptions 1 The flow is steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 Frictional effects associated with sluice gate are negligible. 4 The channel is horizontal.
Sluice gate y1 = 1.1 m
Analysis When frictional effects are negligible and the flow y2 = 0.45 m section is horizontal, the specific energy remains constant, E s1 = E s 2 . Then, V2 V2 V 2 V 2 → 1.1 m + = 0.45 m + y1 + 1 = y 2 + 2 2 2 2 2g 2g 2(9.81 m/s )[(5 m)(1.1 m)] 2(9.81 m/s )[(5 m)(0.45 m)] 2
Solving for the flow rate gives V = 8.806 m 3 /s . The downstream velocity and Froude number are V2 3.914 m/s V V 8.806 m 3 /s V2 = = = = 3.914 m/s and = = 1.86 Fr2 = Ac by 2 (5 m)(0.45 m) gy 2 (9.81 m/s 2 )(0.45 m) Discussion
The actual values will be somewhat different because of frictional effects.
13-105E Solution Water is released from a reservoir through a sluice gate with free outflow. For specified flow depths, the flow rate per unit width and the downstream Froude number are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Analysis For free outflow, we only need the depth ratio y1/a to determine the discharge coefficient (for drowned outflow, we also need to know y2/a and thus the flow depth y2 downstream the gate) y1 5 ft = = 4.55 a 1.1 ft
Sluice gate y1 = 5 ft
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.55. Then the discharge rate becomes
Free outflow
a = 1.1 ft
V = C d ba 2 gy1 = 0.55 (1 ft)(1.1 ft) 2 (32.2 ft/s 2 )(5 ft) = 10.9 ft 3 /s
The specific energy of a fluid remains constant during horizontal flow when the frictional effects are negligible, E s1 = E s 2 . With these approximations, the flow depth past the gate and the Froude number are determined to be E s1 = y1 +
V12 (10.9 ft/s 2 ) 2 V 2 = y1 + = + = 5.074 ft 5 ft 2g 2(32.2 ft/s 2 )[(1 ft)(5 ft)] 2 2 g (by1 ) 2
E s2 = y 2 +
V 22 (10.9 ft/s 2 ) 2 V 2 = y2 + = E → y + = 5.074 ft s 1 2 2g 2 g (by 2 ) 2 2(32.2 ft/s 2 )[(1 ft)( y 2 )] 2
Solution yields y2 = 0.643 ft as the physically meaningful root (positive and less than 5 ft). Then, V2 16.9 ft/s V V 10.9 ft 3 /s and = = 3.71 Fr2 = V2 = = = = 16.9 ft/s Ac by 2 (1 ft)(0.643 ft) gy 2 (32.2 ft/s 2 )(0.643 ft) Discussion In actual gates some frictional losses are unavoidable, and thus the actual velocity and Froude number downstream will be lower.
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Chapter 13 Open-Channel Flow 13-106E Solution Water is released from a reservoir through a drowned sluice gate into an open channel. For specified flow depths, the rate of discharge is to be determined. Sluice gate Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible.
The depth ratio y1/a and the contraction coefficient y2/a are
Analysis
y1 5 ft = = 4.55 a 1.1 ft
and
y 2 3.3 ft = =3 a 1.1 ft
y1 = 5 ft
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.44. Then the discharge rate becomes
y2 = 2 ft
a = 1.1 ft
V = C d ba 2 gy1 = 0.44 (1 ft)(1.1 ft) 2 (32.2 ft/s 2 )(5 ft) = 8.69 ft 3 /s
Then the Froude number downstream the gate becomes V2 =
V V 8.69 ft 3 /s = = = 2.63 ft/s Ac by 2 (1 ft)(3.3 ft)
Discussion
→
Fr2 =
V2 gy 2
=
2.63 ft/s (32.2 ft/s 2 )(3.3 ft)
= 0.255
Note that the flow past the gate becomes subcritical when the outflow is drowned.
13-107 Solution Water is released from a lake through a drowned sluice gate into an open channel. For specified flow depths, the rate of discharge through the gate is to be determined. Sluice gate Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Analysis are
The depth ratio y1/a and the contraction coefficient y2/a
y1 6m = = 10 a 0.6 m
and
y1 = 6 m
y2 3m = =5 a 0.6 m
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.48. Then the discharge rate becomes
y2 = 3 m
a = 0.6 m
V = C d ba 2 gy1 = 0.48 (5 m)(0.6 m) 2 (9.81 m/s 2 )(6 m) = 15.6 m 3 /s
Discussion Note that the use of the discharge coefficient enables us to determine the flow rate through sluice gates by measuring 3 flow depths only.
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Chapter 13 Open-Channel Flow 13-108E Solution Water discharged through a sluice gate undergoes a hydraulic jump. The flow depth and velocities before and after the jump and the fraction of mechanical energy dissipated are to be determined. Assumptions 1 The flow is steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 Frictional effects associated with sluice gate are negligible. 4 The channel is horizontal.
Sluice gate
Analysis For free outflow, we only need the depth ratio y1/a to determine the discharge coefficient,
Hydraulic jump
y1 = 8 ft
y1 8 ft = =8 a 1 ft
a =1 ft
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.58. Then the discharge rate becomes
y3 y2
V = C d ba 2 gy1 = 0.58 (1 ft)(1 ft) 2 (32.2 ft/s 2 )(8 ft) = 13.16 ft 3 /s
The specific energy of a fluid remains constant during horizontal flow when the frictional effects are negligible, E s1 = E s 2 . With these approximations, the flow depth past the gate and the Froude number are determined to be V2 (13.16 ft 3 /s) 2 V 2 E s1 = y1 + 1 = y1 + = 8 ft + = 8.042 ft 2g 2 g (by1 ) 2 2(32.2 ft/s 2 )[(1 ft)(8 ft)] 2 V2 (13.16 ft 3 /s) 2 V 2 E s2 = y 2 + 2 = y 2 + = E → y + = 8.042 ft s1 2 2g 2 g (by 2 ) 2 2(32.2 ft/s 2 )[(1 ft)( y 2 )] 2 It gives y2 = 0.601 ft as the physically meaningful root (positive and less than 8 ft). Then, V V 13.16 ft 3 /s = = = 21.9 ft/s Ac by 2 (1 ft)(0.601 ft) V2 21.9 ft/s Fr2 = = = 4.97 gy 2 (32.2 ft/s 2 )(0.601 ft)
V2 =
Then the flow depth and velocity after the jump (state 3) become y 3 = 0.5 y 2 ⎛⎜ − 1 + 1 + 8Fr22 ⎞⎟ = 0.5(0.601 ft)⎛⎜ − 1 + 1 + 8 × 4.97 2 ⎞⎟ = 3.94 ft ⎠ ⎝ ⎠ ⎝ y2 0.601 ft V2 = (21.9 ft/s) = 3.34 ft/s y3 3.94 ft The head loss and the fraction of mechanical energy dissipated during the jump are V3 =
hL = y 2 − y 3 +
V 22 − V32 (21.9 ft/s) 2 - (3.34 ft/s) 2 = (0.601 ft) - (3.94 ft) + = 3.93 ft 2g 2(32.2 ft/s 2 )
Dissipation ratio =
Discussion
hL hL 3.93 ft = = = 0.488 E s 2 y 2 (1 + Fr22 / 2) (0.601 ft)(1 + 4.97 2 / 2)
Note that almost half of the mechanical energy of the fluid is dissipated during hydraulic jump.
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Chapter 13 Open-Channel Flow 13-109 Solution The flow rate of water in an open channel is to be measured with a sharp-crested triangular weir. For a given flow depth upstream the weir, the flow rate is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Properties
The weir discharge coefficient is given to be 0.60.
Analysis
The discharge rate of water is determined directly from 8 ⎛θ ⎞ V = C wd ,tri tan⎜ ⎟ 2 g H 5 / 2 15 ⎝ 2 ⎠
Upstream free surface
1m
60°
Weir plate
0.5 m
where Cwd = 0.60, θ = 60°, and H = 1 m. Substituting,
3m
8 ⎛ 60° ⎞ 2 5/2 3 V = (0.60 ) tan⎜ ⎟ 2 (9.81 m/s ) (1 m) = 0.818 m /s 15 ⎝ 2 ⎠
Discussion Note that the use of the discharge coefficient enables us to determine the flow rate in a channel by measuring a single flow depth. Triangular weirs are best-suited to measure low discharge rates as they are more accurate than the other weirs for small heads.
13-110 Solution The flow rate of water in an open channel is to be measured with a sharp-crested triangular weir. For a given flow depth upstream the weir, the flow rate is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Properties
The weir discharge coefficient is given to be 0.60.
Analysis
The discharge rate of water is determined directly from V = C wd ,tri
8 ⎛θ ⎞ tan⎜ ⎟ 2 g H 5 / 2 15 ⎝ 2 ⎠
Upstream free surface
0.7 m
60°
Weir plate
0.5 m
where Cwd = 0.60, θ = 60°, and H = 0.7 m. Substituting,
3m
8 ⎛ 60° ⎞ 2 5/2 3 V = (0.60 ) tan⎜ ⎟ 2 (9.81 m/s ) (0.7 m) = 0.335 m /s 15 2 ⎝ ⎠ Discussion Note that the use of the discharge coefficient enables us to determine the flow rate in a channel by measuring a single flow depth. Triangular weirs are best-suited to measure low discharge rates as they are more accurate than the other weirs for small heads.
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Chapter 13 Open-Channel Flow 13-111 Solution The notch angle of a sharp-crested triangular weir used to measure the discharge rate of water from a lake is reduced by half. The percent reduction in the discharge rate is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The water depth in the lake and the weir discharge coefficient remain unchanged.
H
The discharge rate through a triangular weir is given as 8 ⎛θ ⎞ V = C wd ,tri tan⎜ ⎟ 2 g H 5 / 2 15 ⎝ 2 ⎠
θ°
Analysis
Weir plate
Pw
Therefore, the discharge rate is proportional to the tangent of the half notch angle, and the ratio of discharge rates is calculated to be V tan(50° / 2) = 0.391 V = 50° = tan( 100° / 2) V100° When the notch angle is reduced by half, the discharge rate drops to 39.1% of the original level. Therefore, the percent reduction in the discharge rate is Percent reduction = 1-0.391 =0.609= 60.9% Discussion Note that triangular weirs with small notch angles can be used to measure small discharge rates while weirs with large notch angles can be used to measure for large discharge rates.
13-112 Solution The flow rate in an open channel is to be measured using a broad-crested rectangular weir. For a measured value of flow depth upstream, the flow rate is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis The weir head is H = y1 − Pw = 1.6 − 1.0 = 0.6 m . The discharge coefficient of the weir is 0.65 0.65 = = 0.5139 C wd ,broad = 1 + H / Pw 1 + (0.6 m)/(1.0 m)
Discharge
y1 = 1.6 m Pw = 1 m Broad-crested weir
Then the water flow rate through the channel becomes ⎛2⎞ Vrec = Cwd ,broad b g ⎜ ⎟ ⎝3⎠
3/ 2
H 3 / 2 = ( 0.5139 )( 5 m )( 2/3)
3/2
9.81 m/s 2 ( 0.6 m )
3/ 2
= 2.04 m3 /s
The minimum flow depth above the weir is the critical depth, which is determined from ⎛ V 2 y min = y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (2.04 m 3 /s) 2 = ⎜⎜ 2 2 ⎝ (9.81 m/s )(5 m)
⎞ ⎟ ⎟ ⎠
1/ 3
= 0.257 m
Discussion The upstream velocity and the upstream velocity head are V12 (0.255 m/s) 2 V 2.04 m 3 /s V1 = = = 0.255 m/s and = = 0.0033 m by1 (5 m)(1.6 m) 2 g 2(9.81 m/s 2 ) This is 0.3% of the weir head, which is negligible. When the upstream velocity head is considered (by replacing H in the flow rate relation by H + V12 / 2 g ), the flow rate becomes 2.05 m3/s, which is practically identical to the value determined above.
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Chapter 13 Open-Channel Flow 13-113 Solution The flow rate in an open channel is to be measured using a broad-crested rectangular weir. For a measured value of flow depth upstream, the flow rate is to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis
The weir head is H = y1 − Pw = 2.2 − 1.0 = 1.2 m . The discharge coefficient of the weir is
C wd , broad =
0.65
=
1 + H / Pw
0.65 1 + (1.2 m)/(1.0 m)
= 0.4382
Then the water flow rate through the channel becomes Vrec
⎛2⎞ = C wd ,broad b g ⎜ ⎟ ⎝ 3⎠
3/ 2
= (0.4382)(5 m)(2/3)
H 3/2
Discharge
y1 = 2.2 m
3/ 2
Pw = 1 m 2
9.81 m/s (1.2 m )
3/ 2
Broad-crested weir
= 4.91 m /s 3
The minimum flow depth above the weir is the critical depth, which is determined from ⎛ V 2 ⎞ y min = y c = ⎜⎜ 2 ⎟⎟ ⎝ gb ⎠
1/ 3
⎛ (4.91 m 3 /s) 2 = ⎜⎜ 2 2 ⎝ (9.81 m/s )(5 m)
⎞ ⎟ ⎟ ⎠
1/ 3
= 0.462 m
Discussion The upstream velocity and the upstream velocity head are V12 (0.446 m/s) 2 V 4.91 m 3 /s V1 = = = 0.446 m/s and = = 0.010 m by1 (5 m)(2.2 m) 2 g 2(9.81 m/s 2 ) This is 0.8% of the weir head, which is negligible. When the upstream velocity head is considered (by replacing H in the flow rate relation by H + V12 / 2 g , the flow rate becomes 4.97 m3/s, which is practically identical to the value determined above.
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Chapter 13 Open-Channel Flow 13-114 Solution The flow rate in an open channel is measured using a broad-crested rectangular weir. For a measured value of minimum flow depth over the weir, the flow rate and the upstream flow dept are to be determined. Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis The flow depth over the reaches its minimum value when the flow becomes critical. Therefore, the measured minimum depth is the critical depth yc. Then the flow rate is determined from the critical depth relation to be y min
⎛ V 2 = y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
→ V =
y c3 gb 2 = (0.50 m) 3 (9.81 m/s 2 )(1 m) 2 = 1.11 m 3 /s
This is the flow rate per m width of the channel since we have taken b = 1 m. Disregarding the upstream velocity head and noting that the discharge coefficient of the weir is C wd , broad = 0.65 / 1 + H / Pw , the flow rate for a broad-crested weir can be expressed as Vrec =
0.65
⎛2⎞ b g⎜ ⎟ ⎝ 3⎠ 1 + H/Pw
3/ 2
H 3/ 2
Discharge
Substituting,
y1
1.11 m 3 /s =
0.65 m 1 + H/(0.8 m)
(1 m)(2/3) 3/2 9.81 m/s 2 H 3 / 2
= 4.91 m 3 /s
yc = 0.50 m Pw = 0.80 m Broad-crested weir
Its solution is H = 1.40 m. Then the flow depth upstream the weir becomes y1 = H + Pw = 1.40 + 0.80 = 2.20 m
The upstream velocity and the upstream velocity head are V12 (0.503 m/s) 2 V 1.11 m 3 /s V1 = = = 0.503 m/s and = = 0.013 m by1 (1 m)(2.2 m) 2 g 2(9.81 m/s 2 ) This is 0.9% of the weir head, which is negligible. When the upstream velocity head is considered (by replacing H in the flow rate relation by H + V12 / 2 g , the flow rate becomes 1.12 m3/s, which is practically identical to the value determined above. Discussion
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Chapter 13 Open-Channel Flow
Review Problems
13-115 Solution Water flows uniformly in a trapezoidal channel. For a given flow depth, it is to be determined whether the flow is subcritical or supercritical. Assumptions
The flow is uniform.
Analysis
The flow area and the average velocity are
y = 0.6 m (b + b + 2 y / tan θ ) [4 + 4 + 2(0.60 m)/tan45°] m = (0.60 m) = 2.76 m 2 θ =45° 2 2 b=4m 18 m 3 s V = = 6 . 522 m/s V= Ac 2.76 m 2 When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, 2.76 m 2 Ac Ac = = = 0.5308 m y = yh = Top width b + 2 y / tan θ (4 + 2 × 0.60 / tan 45°) m 6.522 m/s V Then the Froude number becomes Fr = = = 2.86 , which is greater than 1. gy (9.81 m/s 2 )(0.5308 m) Ac = y
Therefore, the flow is supercritical. Discussion The analysis is approximate since the edge effects are significant here compared to a wide rectangular channel, and thus the results should be interpreted accordingly.
13-116 Solution determined. Assumptions
Water flows in a rectangular channel. The flow depth below which the flow is supercritical is to be 8 m3/s
The flow is uniform.
Analysis The flow depth below which the flow is super critical is the critical depth yc determined from
⎛ V 2 y c = ⎜⎜ 2 ⎝ gb
⎞ ⎟ ⎟ ⎠
1/ 3
⎛ (8 m 3 /s) 2 =⎜ ⎜ (9.81 m/s 2 )(2 m) 2 ⎝
⎞ ⎟ ⎟ ⎠
1/3
y
= 1.18 m 2m
Therefore, flow is supercritical for y < 1.18 m. Discussion Note that a flow is more likely to exist as supercritical when the flow depth is low and thus the flow velocity is high.
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Chapter 13 Open-Channel Flow 13-117 Solution Water flows in a canal at a specified average velocity. For various flow depths, it is to be determined whether the flow is subcritical or supercritical. Assumptions
The flow is uniform.
Analysis
For each depth, we determine the Froude number and compare it to the critical value of 1:
(a) y = 0.2 m:
Fr =
V gy
=
4 m/s (9.81 m/s 2 ) (0.2 m)
= 2.86 > 1
4 m/s
which is greater than 1. Therefore, the flow is supercritical. (b) y = 2 m:
Fr =
V gy
=
4 m/s (9.81 m/s 2 ) (2 m)
= 0.903 < 1
y
which is less than 1. Therefore, the flow is subcritical. (c) y = 1.63 m:
Fr =
V gy
=
4 m/s (9.81 m/s 2 ) (1.63 m)
=1
b
which is equal to 1. Therefore, the flow is critical. Discussion Note that a flow is more likely to exist as supercritical when the flow depth is low and thus the flow velocity is high. Also, the type of flow can be determined easily by checking Froude number.
13-118 Solution The flow of water in a rectangular channel is considered. For a given flow depth and bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
The Manning coefficient is given to be n = 0.012 (Table 13-1).
Analysis
The flow area, wetted perimeter, and hydraulic radius of the channel are
Ac = by = (1.5 m)(0.9 m) = 1.35 m 2
Rh =
p = 1.5 m + 2 × 0.9 m = 3.3 m
Ac 1.35 m 2 = = 0.4091 m p 3.3 m y = 0.9 m
Bottom slope of the channel is S 0 = tan 0.6 D = 0.01047
Then the flow rate can be determined from Manning’s equation to be
b = 1.5 m
1/3
a 1m / s V = Ac R h2 / 3 S 01 / 2 = (1.35 m 2 )(0.4091 m) 2/3 (0.01047) 1 / 2 = 6.34 m 3 s n 0.012
Discussion
Note that the flow rate in a given channel is a strong function of the bottom slope.
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Chapter 13 Open-Channel Flow
13-119
Solution The flow of water in a rectangular channel is considered. The effect of bottom slope on the flow rate is to be investigated as the bottom angle varies from 0.5 to 10°. Assumptions
1 The flow is steady and uniform. 2 Roughness coefficient is constant along the channel.
Properties
Manning coefficient for an open channel made of finished concrete is n = 0.012 (Table 13-1).
Analysis
The EES Equations window is printed below, along with the tabulated and plotted results.
a=1 b=5 Vdot = 12 "m3/s" n=0.012
y
s=tan(teta) Ac=b*y p=b+2*y Rh=Ac/p Vdot=(a/n)*Ac*Rh^(2/3)*SQRT(s)
θ°
0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0 Discussion
Flow depth, y, m 0.533 0.427 0.375 0.343 0.320 0.302 0.287 0.276 0.266 0.257 0.250 0.243 0.237 0.231 0.226 0.222 0.218 0.214 0.210 0.207
0.55
0.5
0.45
0.4
Flow depth, y (m)
Bottom angle,
b=5m
0.35
0.3
0.25
0.2 0
2
4
6
8
10
Bottom angle, θ
Note that the flow depth decreases as the bottom angle increases, as expected.
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Chapter 13 Open-Channel Flow
13-120
Solution The flow of water in a trapezoidal channel is considered. The effect of bottom slope on the flow rate is to be investigated as the bottom angle varies from 0.5 to 10°. Assumptions
1 The flow is steady and uniform. 2 Roughness coefficient is constant along the channel.
Properties
Manning coefficient for an open channel made of finished concrete is n = 0.012 (Table 13-1).
Analysis
The EES Equations window is printed below, along with the tabulated and plotted results.
a=1 b=5 Vdot = 12 "m3/s" n=0.012
y
s=tan(teta) Ac=y*(b+y/tan(45)) p=b+2*y/sin(45) Rh=Ac/p Vdot=(a/n)*Ac*Rh^(2/3)*SQRT(s)
θ°
0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0 Discussion
Flow depth, y, m 0.496 0.403 0.357 0.327 0.306 0.290 0.276 0.266 0.256 0.248 0.241 0.235 0.229 0.224 0.220 0.215 0.211 0.208 0.204 0.201
b=5m
0.5
0.45
0.4
Flow depth y (m)
Bottom angle,
θ =45°
0.35
0.3
0.25
0.2 0
2
4
6
8
10
Bottom Angle, θ
As expected, flow depth decreases with increasing bottom angle, but the relationship is far from linear.
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PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-121
Solution The flow of water in a trapezoidal channel is considered. For a given flow depth and bottom slope, the flow rate is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
The Manning coefficient for a brick-lined open channel is n = 0.015 (Table 13-1).
The flow area, wetted perimeter, and hydraulic radius of the channel are 2m ⎞ y ⎞ ⎛ ⎛ Ac = y⎜ b + ⎟ = 14.93 m 2 ⎟ = (2 m) ⎜ 4 m + tan θ ⎠ tan30 ° ⎝ ⎝ ⎠
Analysis
p =b+ Rh =
2 (2 m) 2y =4m+ = 12 m sin30° sin θ
y=2m
θ = 30°
30°
2
Ac 14.93 m = = 1.244 m p 12 m
b=4
Bottom slope of the channel is So = 0.001.Then the flow rate can be determined from Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 = (14.93 m 2 )(1.244 m) 2/3 (0.001) 1 / 2 = 36.4 m 3 s n 0.015
Note that the flow rate in a given channel is a strong function of the bottom slope.
Discussion
13-122 Solution The flow of water in a circular open channel is considered. For given flow depth and flow rate, the elevation drop per km length is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Properties
The Manning coefficient for the steel channel is given to be n = 0.012.
Analysis
The flow area, wetted perimeter, and hydraulic radius of the channel are
y − R 1.5 − 1 π 2π = = 0.5 → α = 60° = 60 = R 1 360 3 π 2π θ = π −α = π − = = 120° 3 3
cos α =
2
2
Ac = R (θ − sin θ cos θ ) = (1 m) [2π / 3 − sin( 2π / 3) cos(2π / 3)] = 2.527 m Rh =
y-R 2
α
R=1m
θ
Ac θ − sin θ cos θ 2π / 3 − sin( 2π / 3) cos(2π / 3) = R= (1 m) = 0.6034 m p 2θ 2 × 2π / 3
y= 1.5 m
Substituting the given quantities into Manning’s equation, 1 m 1/3 / s a (2.527 m 2 )(0.6034 m) 2/3 S 01 / 2 V = Ac R h2 / 3 S 01 / 2 → 12 m 3 /s = 0.012 n It gives the slope to be S0 = 0.00637. Therefore, the elevation drop Δz across a pipe length of L = 1 km must be Δz = S 0 L = 0.00637(1000 m) = 6.37 m
Discussion Note that when transporting water through a region of fixed elevation drop, the only way to increase the flow rate is to use a channel with a larger cross-section.
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Chapter 13 Open-Channel Flow 13-123 Solution The flow of water through a V-shaped open channel is considered. The angle θ the channel makes from the horizontal is to be determined for the case of most efficient flow. Assumptions
1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness coefficient is constant.
Analysis
We let the length of the sidewall of the channel be x. From trigonometry,
y b → y = x sin θ cos θ = → b = x cos θ x x Then the cross-sectional area and the perimeter of the flow section become 2 Ac x2 sin 2θ Ac = by = x cos θ sin θ = → x= 2 sin 2θ
b
sin θ =
x
θ
y
θ
2 Ac → p = 2 2 Ac (sin 2θ ) −1 / 2 sin 2θ Now we apply the criterion that the best hydraulic cross-section for an open channel is the one with the minimum wetted perimeter for a given cross-section. Taking the derivative of p with respect to θ while holding Ac constant gives dp d [(sin 2θ ) −1 / 2 ] d [(sin 2θ ) −1 / 2 ] d (sin 2θ ) −3 = 2 2 Ac = 2 2 Ac = 2 2 Ac 2 cos 2θ dθ dθ d (sin 2θ ) dθ 2(sin 2θ ) 3 / 2 p = 2x = 2
Setting dp/dθ = 0 gives cos 2θ = 0 , which is satisfied when 2θ = 90°. Therefore, the criterion for the best hydraulic crosssection for a triangular channel is determined to be θ = 45°. Discussion
The procedure used here can be used to determine the best hydraulic cross-section for any geometric shape.
13-124E Solution Water is to be transported in a rectangular channel at a specified rate. The dimensions for the best crosssection if the channel is made of unfinished concrete are to be determined. Assumptions
1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The roughness coefficient is constant.
Properties
The Manning coefficient is n = 0.014 for channels made of unfinished concrete (Table 13-1).
Analysis
For best cross-section of a rectangular cross-section, y = b/2. Then Ac = yb = b2/2, and Rh = b/4.
a The flow rate is determined from the Manning equation, V = Ac R h2 / 3 S 01 / 2 . n
(a) Bottom drop of 8 ft per mile: s = (8 ft) /(5280 ft) = 0.001515 200 ft 3 /s =
1.486 ft 1/3 / s 2 (b / 2)(b / 4 ) 2/3 (0.001515) 1 / 2 0.014
y = b/2
Solving the above equation gives b = 7.86 ft, and y = b/2 = 3.93 ft. (b) Bottom drop of 10 ft per mile: s = (10 ft) /(5280 ft) = 0.001894 200 ft 3 /s =
b
1.486 ft 1/3 / s 2 (b / 2)(b / 4 ) 2/3 (0.001894 ) 1 / 2 0.014
Solving the above equation gives b = 7.54 ft, and y = b/2 = 3.77 ft. Discussion The concept of best cross-section is an important consideration in the design of open channels because it directly affects the construction costs.
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Chapter 13 Open-Channel Flow 13-125E Solution Water is to be transported in a trapezoidal channel at a specified rate. The dimensions for the best crosssection if the channel is made of unfinished concrete are to be determined. Assumptions 1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the channel. Properties
The Manning coefficient is n = 0.014 for channels made of unfinished concrete (Table 13-1).
Analysis
For best cross-section of a trapezoidal channel of bottom width b, θ = 60° and y = b 3 / 2 . Then,
Ac = y (b + b cos θ ) = 0.5 3b 2 (1 + cos 60°) = 0.75 3b 2 , p = 3b , and R h =
y 3 = b. 2 4
a The flow rate is determined from the Manning equation, V = Ac R h2 / 3 S 01 / 2 , n
(a) Bottom drop of 8 ft per mile: s = (8 ft) /(5280 ft) = 0.001515
200 ft 3 /s =
1.486 ft 1/3 / s (0.75 3b 2 )( 3b / 4) 2/3 (0.001515) 1 / 2 0.014
y=
b
3 b 2
60°
It gives b = 4.79 ft, and y = 4.15 ft.
b
(b) Bottom drop of 10 ft per mile: s = (10 ft) /(5280 ft) = 0.001894 200 ft 3 /s =
1.486 ft 1/3 / s (0.75 3b 2 )( 3b / 4 ) 2/3 (0.001894 ) 1 / 2 0.014
It gives b = 4.59 ft, and y = 3.98 ft. The concept of best cross-section is an important consideration in the design of open channels because it Discussion directly affects the construction costs.
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Chapter 13 Open-Channel Flow 13-126 Solution Water is flowing through a channel with nonuniform surface properties. The flow rate through the channel and the effective Manning coefficient are to be determined. Assumptions 1 The flow is steady and uniform. 2 The bottom slope is constant. 3 The Manning coefficients do not vary along the channel.
10 m
Clean earth channel 1m Analysis The channel involves n1 = 0.022 two parts with different roughness, and thus it is appropriate to divide the 1m channel into two subsections. The flow rate for each subsection can be 3m determined from the Manning equation, and the total flow rate can be determined by adding them up. The flow area, perimeter, and hydraulic radius for each subsection and the entire channel are: A 6 m2 Subsection 1: Ac1 = 6 m 2 , p1 = 6 m , R h1 = c1 = = 1.00 m p1 6m Subsection 2:
Ac 2 = 10 m 2 ,
p 2 = 11 m ,
Rh2 =
Heavy brush n2 = 0.075
Ac 2 10 m 2 = = 0.909 m p2 11 m
Ac 16 m 2 = = 0.941 m p 17 m Applying the Manning equation to each subsection, the total flow rate through the channel becomes
Entire channel:
Ac = 16 m 2 ,
p = 17 m ,
Rh =
a a A1 R12 / 3 S 01 / 2 + A2 R 22 / 3 S 01 / 2 V = V1 +V 2 = n1 n2 ⎛ (6 m 2 ) (1 m) 2/3 (10 m 2 )(0.909 m) 2/3 = (1 m 1/3 /s) ⎜ + ⎜ 0.022 0.075 ⎝
⎞ ⎟ (tan0.5°)1/2 ⎟ ⎠
= 37.2 m 3 /s
Knowing the total flow rate, the effective Manning coefficient for the entire channel can be determined from the Manning equation to be n eff =
aAc R h2 / 3 S 01 / 2 (1 m 1/3 / s)(16 m 2 )(0.941 m) 2/3 (0.00873)1 / 2 = = 0.0386 V 37.2 m 3 / s
Discussion The effective Manning coefficient neff of the channel turns out to lie between the two n values, as expected. The weighted average of the Manning coefficient of the channel is nave=(n1p1+ n2p2)/p = 0.056, which is quite different than neff. Therefore, using a weighted average Manning coefficient for the entire channel may be tempting, but it would not be so accurate.
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Chapter 13 Open-Channel Flow 13-127 Solution Two identical channels, one rectangular of bottom width b and one circular of diameter D, with identical flow rates, bottom slopes, and surface linings are considered. The relation between b and D is to be determined for the case of the flow height y = b and the circular channel is flowing half full. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Analysis
The cross-sectional area, perimeter, and hydraulic radius of the rectangular channel are
Ac = b 2 ,
p = 3b , and R h =
Ac b 2 b = = p 3b 3
Then using the Manning equation, the flow rate can be expressed as a a ⎛b⎞ Vrec = Ac R h2 / 3 S 01 / 2 = b 2 ⎜ ⎟ n n ⎝ 3⎠
2/3
S 01 / 2 =
D
a 1/ 2 b 8 / 3 S0 n 32 / 3
The corresponding relations for the semi-circular channel are Ac =
πD 2 8
,
p=
πD 2
, and Rh =
Ac D = p 4
and a a D2 ⎛ D ⎞ Vcir = Ac R h2 / 3 S 01 / 2 = π ⎜ ⎟ 8 ⎝4⎠ n n
2/3
S 01 / 2 =
a 1 / 2 πD 8 / 3 S0 n 8× 4 2 / 3
b
Setting the flow rates in the two channels equal to each other Vcir = Vrec gives 8/ 3
8/ 3
8/3
8/ 3
b πD a 1/ 2 b a πD S0 = S 1/ 2 → 2 / 3 = 2/3 2/3 0 n n 8× 4 3 3 8× 42/ 3
Therefore, the desired relation is
→
b ⎛ π3 =⎜ D ⎜⎝ 8 × 4 2 / 3 2/3
⎞ ⎟ ⎟ ⎠
b 3/ 8
= 0.655
b = 0.655D.
Discussion Note that the wetted perimeters in this case are prec = 3b = 2.0D and pcir = πD/2 = 1.57D. Therefore, the semi-circular channel is a more efficient channel than the rectangular one.
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Chapter 13 Open-Channel Flow 13-128 Solution The flow of water through a parabolic notch is considered. A relation is to be developed for the flow rate, and its numerical value is to be calculated. Assumptions 1 The flow is steady. 2 All frictional effects are negligible, and Toricelli’s equation can be used for the velocity. Analysis The notch is parabolic with y = 0 at x = 0, and thus it can be expressed analytically as y = Cx 2 . Using the coordinates of the upper right corner, the value of the constant is determined to be
b = 0.4 m
dy
2x H = 0.5 m
y
C = y / x = H / (b / 2) = 4H / b = 2
2
x
2
4 ( 0.5 m ) / ( 0.4 m ) = 12.5 m-1 . 2
A differential area strip can be expressed as dA = 2 xdy = 2 y / C dy
Noting that the flow velocity is V = 2 g ( H − y ) , the flow rate through this differential area is
(
)
VdA = V 2 y / C dy = 2 g ( H − y ) 2 y / C dy = 2 2 g / C y ( H − y ) dy
Then the flow rate through the entire notch is determined by integration to be V =
∫
VdA = 2 2 g / C
A
∫
H
y =0
y ( H − y ) dy
where H
⎡ ⎛ 2 y − H ⎞⎤ 1 H2 ⎟⎥ = π H 2 y ( H − y ) dy = ⎢ (2 y − H ) Hy − y 2 + Arc tan⎜⎜ 2 ⎢4 y =0 8 16 ⎜ 2 Hy − y ⎟⎟⎥ ⎝ ⎠⎦ 0 ⎣ Then the expression for the volume flow rate and its numerical value become
∫
H
π V = 8 Discussion K = Cd
π 8
2g 2 π H = C 8
2(9.81 m/s 2 ) 12.5 m -1
H 2 = (0.492 m/s) H 2 = (0.492 m/s)(0.5 m) 2 = 0.123 m 3 /s
Note that a general flow rate equation for parabolic notch would be in the form of V = KH 2 , where 2g and Cd is the discharge coefficient whose value is determined experimentally to account for nonideal C
effects.
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Chapter 13 Open-Channel Flow
13-129
Solution The flow of water through a parabolic notch is considered. A relation is to be developed for the flow rate, and its numerical value is to be calculated. 0.07
Assumptions 1 The flow is steady. 2 All frictional effects are negligible, and Toricelli’s equation can be used for the velocity.
0.06 0.05
b = 2y tan(θ/2)
0.04 3
V, m /s
d
H = 25 y
0.03 0.02
θ 0.01 0 0
Consider a differential strip area shown Analysis on the sketch. It can be expressed as
10
dA = bdy = 2 y tan(θ / 2)dy
20
30
40
θ, °
50
60
70
80
90
Noting that the flow velocity is V = 2 g ( H − y ) , the flow rate through this differential area is VdA = V (2 y tan(θ / 2)dy ) = 2 g ( H − y ) 2 y tan(θ / 2)dy = 2 2 g tan(θ / 2) y H − y dy
Then the flow rate through the entire notch is determined by integration to be V =
∫
A
VdA = 2 2 g tan(θ / 2)
∫
H
y =0
y H − y dy
where
∫
H
y =0
2 2 y H − y dy = ⎡⎢− y 5 / 2 + Hy 3 / 2 ⎤⎥ 3 ⎦ ⎣ 5
H
= 0
4 5/ 2 H 15
Then the expression for the volume flow rate and its numerical value become 8 2g 8 2(9.81 m/s 2 ) tan(θ / 2) H 5 / 2 = V = tan(θ / 2)(0.25) 5 / 2 = 0.07382 tan(θ / 2) (m3/s) 15 15
θ = 25°: V = 0.07382tan ( 25° / 2 ) = 0.0164 m3 /s θ = 40°: V = 0.07382tan ( 40° / 2 ) = 0.0269 m3 /s θ = 60°: V = 0.07382tan ( 60° / 2 ) = 0.0426 m3 /s θ = 75°: V = 0.07382tan ( 75° / 2 ) = 0.0566 m3 /s These results are plotted, using EES. Discussion
Note that a general flow rate equation for the V-notch would be in the form of V = K tan(θ / 2) H 5 / 2 ,
where K = C d 8 2 g / 15 and Cd is the discharge coefficient whose value is determined experimentally to account for nonideal effects.
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Chapter 13 Open-Channel Flow 13-130 Solution Water flows uniformly half-full in a circular channel. For specified flow rate and bottom slope, the Manning coefficient is to be determined. Assumptions channel.
1 The flow is steady and uniform. 2 Bottom slope is constant. 3 Roughness coefficient is constant along the
Analysis
The flow area, wetted perimeter, and hydraulic radius of the channel are 2 πR 2 π (0.6 m) 2
Ac =
= = 0.5655 m 2 2 2πR 2π (0.6 m) p= = = 1.885 m 2 2 A πR 2 / 2 R 0.60 m Rh = c = = = = 0.30 m P 2 2 πR
R = 0.6 m
θ = π/2
Then the Manning coefficient can be determined from Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 → 1.25 m 3 /s = (0.5655 m 2 )(0.30 m) 2/3 (0.004)1 / 2 n n It gives the Manning coefficient to be n = 0.013
When calculating the Froude number, the hydraulic depth should be used rather than the maximum depth or the hydraulic radius. For a non-rectangular channel, hydraulic depth is defined as the ratio of the flow area to top width, Ac πR 2 / 2 πR π (0.6 m) yh = = = = = 0.4712 m Top width 2R 4 4 1.25 m 3 s V = = 2.210 m s V= Ac 0.5655 m 2 V 2.21m/s Fr = = = 1.03 gy (9.81 m/s 2 )(0.4721 m) which is greater than 1. Therefore, the flow is supercritical. Discussion
It appears that this channel is made of cast iron or unplaned wood .
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Chapter 13 Open-Channel Flow 13-131 Solution determined.
Water flowing in a horizontal open channel encounters a bump. Flow properties over the bump are to be
Assumptions 1 The flow is steady. 2 Frictional effects are negligible so that there is no dissipation of mechanical energy. 3 The channel is sufficiently wide so that the end effects are negligible.
The upstream Froude number and the critical depth are
Analysis Fr1 =
V1 gy1
1.25 m/s
=
(9.81 m 2 /s)( 1.8 m)
y1=1.8 =0.80mm =1.2 m/s VV11=1.25 m/s
y2 V2 20z = cm0.15 m Δ
= 0.297 1/ 3
1/ 3
1/ 3 1/ 3 ⎛ (by V ) 2 ⎞ ⎛ y 2V 2 ⎞ ⎛ V 2 ⎞ ⎛ (1.8 m) 2 (1.25 m / s) 2 ⎞ ⎟ = 0.802 m y c = ⎜⎜ 2 ⎟⎟ = ⎜⎜ 1 21 ⎟⎟ = ⎜⎜ 1 1 ⎟⎟ = ⎜⎜ ⎟ 9.81 m/s 2 ⎝ gb ⎠ ⎝ ⎠ ⎝ gb ⎠ ⎝ g ⎠ The upstream flow is subcritical since Fr < 1, and the flow depth decreases over the bump. The upstream, over the bump, and critical specific energy are
E s1 = y1 +
V12 (1.25 m/s) 2 = (1.80 m) + = 1.88 m 2g 2(9.81 m/s 2 )
The flow depth over the bump can be determined from y 23 − ( E s1 − Δz b ) y 22 +
V12 2 y1 = 0 2g
→
y 23 − (1.88 − 0.20 m) y 22 +
(1.25 m/s) 2 2(9.81 m/s 2 )
(1.80 m) 2 = 0
Using an equation solver, the physically meaningful root of this equation is determined to be y2 = 1.576 m. Then, V2 = Fr2 =
y1 1.8 m V1 = (1.25 m/s ) = 1.43 m/s y2 1.576 m V2 gy 2
=
1.428 m/s (9.81 m/s 2 )( 1.576 m)
= 0.363
Discussion The actual values may be somewhat different than those given above because of the frictional effects that are neglected in the analysis.
13-66
PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 13 Open-Channel Flow 13-132 Solution Water flowing in a horizontal open channel encounters a bump. The bump height for which the flow over the bump is critical is to be determined. Assumptions 1 The flow is steady. 2 Frictional effects are negligible so that there is no dissipation of mechanical energy. 3 The channel is sufficiently wide so that the end effects are negligible.
The upstream Froude number and the critical depth are
Analysis Fr1 =
V1 gy1
=
1.25 m/s (9.81 m 2 /s)( 1.8 m)
y1=1.8 =0.80mm =1.2 m/s VV11=1.25 m/s
y2 V2 20z = cm0.15 m Δ
= 0.297 1/ 3
1/ 3
1/ 3 1/ 3 ⎛ (by1V1 ) 2 ⎞ ⎛ y12V12 ⎞ ⎛ V 2 ⎞ ⎛ (1.8 m) 2 (1.25 m / s) 2 ⎞ ⎜ ⎟ ⎜ ⎟ ⎜ ⎟ ⎜ ⎟ yc = ⎜ 2 ⎟ = ⎜ 2 ⎟ = 0.802 m ⎟ =⎜ g ⎟ =⎜ 9.81 m/s 2 ⎝ gb ⎠ ⎝ ⎠ ⎝ gb ⎠ ⎝ ⎠ The upstream flow is subcritical since Fr < 1, and the flow depth decreases over the bump. The upstream specific energy is
E s1 = y1 +
V12 (1.25 m/s) 2 = (1.80 m) + = 1.88 m 2g 2(9.81 m/s 2 )
Noting that the flow over the bump is critical and that E s 2 = E s1 − Δz b , 3 3 E s 2 = E c = yc = (0.802 m) = 1.20 m 2 2 and Δz b = E s1 − E s 2 = 1.88 − 1.20 = 0.68 m Discussion
If a higher bump is used, the flow will remain critical but the flow rate will decrease (the choking effect).
13-133 Solution Water flow through a wide rectangular channel undergoing a hydraulic jump is considered. It is to be shown that the ratio of the Froude numbers before and after the jump can be expressed in terms of flow depths y1 and y2
before and after the jump, respectively, as Fr1 / Fr2 = Assumptions Analysis
( y 2 / y1 )3 .
1 The flow is steady. 2 The channel is sufficiently wide so that the end effects are negligible.
The Froude number for a wide channel of width b and flow depth y is given as V / by V V V Fr = = = = gy gy by gy b gy 3
Expressing the Froude number before and after the jump and taking their ratio gives V/ ⎛⎜ b gy13 ⎞⎟ Fr1 ⎝ ⎠ = = Fr2 V/ ⎛⎜ b gy 3 ⎞⎟ 2 ⎝ ⎠ which is the desired result.
Discussion
gy 23 gy13
⎛y ⎞ = ⎜⎜ 2 ⎟⎟ ⎝ y1 ⎠
3
V2 FIGURE 13-51 y2 Control volume that encloses the y1 hydraulic jump.
V1
(1)
(2)
Using the momentum equation, other relations such as y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ can also be developed. ⎝ ⎠
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Chapter 13 Open-Channel Flow 13-134 Solution A sluice gate with free outflow is used to control the flow rate of water. For specified flow depths, the flow rate per unit width and the downstream flow depth and velocity are to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. Analysis For free outflow, we only need the depth ratio y1/a to determine the discharge coefficient (for drowned outflow, we also need to know y2/a and thus the flow depth y2 downstream the gate), y1 1.8 m = =6 a 0.30 m
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.57. Then the discharge rate per m width becomes
Sluice gate y1 = 1.8 m
Free outflow
a = 30 cm
V = C d ba 2 gy1 = 0.57 (1 m)(0.30 m) 2 (9.81 m/s 2 )(1.8 m) = 1.02 m 3 /s
The specific energy of a fluid remains constant during horizontal flow when the frictional effects are negligible, E s1 = E s 2 . With these approximations, the flow depth and velocity past the gate become V2 (1.02 m/s 2 ) 2 V 2 E s1 = y1 + 1 = y1 + = + = 1.816 m 1 . 8 m 2g 2(9.81 m/s 2 )[(1 m)(1.8 m)] 2 2 g (by1 ) 2 V2 (1.02 m/s 2 ) 2 V 2 y + = 1.816 m E s2 = y 2 + 2 = y 2 + = E → 2 s1 2g 2 g (by 2 ) 2 2(9.81 m/s 2 )[(1 m)( y 2 )] 2 It gives y2 = 0.179 m for flow depth as the physically meaningful root (positive and less than 1.8 m). Also, V2 =
Discussion lower.
1.02 m 3 /s V V = = = 5.67 m/s Ac by 2 (1 m)(0.179 m) In actual gates some frictional losses are unavoidable, and thus the actual velocity downstream will be
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Chapter 13 Open-Channel Flow 13-135 Solution Water at a specified depth and velocity undergoes a hydraulic jump. The fraction of mechanical energy dissipated is to be determined. Assumptions 1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The channel is horizontal. hL Energy Analysis The Froude number before the hydraulic jump is line Fr1 =
V1 gy1
=
8 m/s (9.81 m/s 2 )(0.45 m)
= 3.808
V1 = 8 m/s V2 FIGURE 13-51 y2 Control volume that encloses the y1 = 0.45 m hydraulic jump.
which is greater than 1. Therefore, the flow is indeed supercritical before the jump. The flow depth, velocity, and Froude number after the jump are
(1)
(2)
y 2 = 0.5 y1 ⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ = 0.5(0.45 m)⎛⎜ − 1 + 1 + 8 × 3.808 2 ⎞⎟ = 2.209 m ⎝ ⎠ ⎝ ⎠ V2 =
y1 0.45 m V1 = (8 m/s ) = 1.630 m/s y2 2.209 m
Fr2 =
V2 gy 2
=
1.630 m/s (9.81 m/s 2 )(2.209 m)
= 0.350
The head loss and the fraction of mechanical energy dissipated during the jump are h L = y1 − y 2 +
V12 − V 22 (8 m/s) 2 - (1.63 m/s) 2 = (0.45 m) - (2.209 m) + = 1.368 m 2g 2(9.81 m/s 2 )
hL hL 1.368 m = = = 0.369 2 E s1 y1 (1 + Fr1 / 2) (0.45 m)(1 + 3.808 2 / 2) Note that almost over one-third of the mechanical energy of the fluid is dissipated during hydraulic jump.
Dissipation ratio =
Discussion
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Chapter 13 Open-Channel Flow 13-136 Solution The flow depth and average velocity of water after a hydraulic jump together with approach velocity to sluice gate are given. The flow rate per m width, the flow depths before and after the gate, and the energy dissipation ratio are to be determined.
V1 = 1.25 m/s y1
Sluice gate y3 = 3 m V3 = 4 m/s
y2
Assumptions negligible.
1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are
Analysis jump is
The flow rate per m width of channel, flow depth before the sluice gate, and the Froude number after the
V = V3 Ac 3 = V3 by 3 = (4 m/s)(1 m)(3 m) = 12 m 3 /s
y1 =
V3 4 m/s y3 = (3 m ) = 9.60 m V1 1.25 m/s V3
Fr3 =
gy 3
=
4 m/s (9.81 m/s 2 )(3 m)
= 0.7373
The flow dept, velocity, and Froude number before the jump are
)
(
)
(
y2 = 0.5 y3 −1 + 1 + 8Fr32 = 0.5(3 m) −1 + 1 + 8 × 0.73732 = 1.969 m ≅ 1.97 m V2 =
y3 3m V3 = (4 m/s ) = 6.094 m/s y2 1.969 m
Fr2 =
V2 gy 2
=
6.094 m/s (9.81 m/s 2 )(1.969 m)
= 1.387
which is greater than 1, and thus the flow before the jump is indeed supercritical. The head loss and the fraction of mechanical energy dissipated during hydraulic jump are hL = y 2 − y 3 +
V 22 − V32 (6.094 m/s) 2 - (4 m/s) 2 = (1.969 m) - (3 m) + = 0.0463 m 2g 2(9.81 m/s 2 )
Dissipation ratio =
Discussion
hL hL 0.0463 m = = = 0.0120 2 E s 2 y 2 (1 + Fr2 / 2) (1.969 m)(1 + 1.387 2 / 2)
Note that this is a “mild” hydraulic jump, and only 1.2% of the mechanical energy is wasted.
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Chapter 13 Open-Channel Flow 13-137 Solution The flow depth and average velocity of water after a hydraulic jump together with approach velocity to sluice gate are given. The flow rate per m width, the flow depths before and after the gate, and the energy dissipation ratio are to be determined.
V1 = 1.25 m/s y1
Sluice gate y3 = 3 m V3 = 2 m/s
y2
Assumptions negligible.
1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are
Analysis jump is
The flow rate per m width of channel, flow depth before the sluice gate, and the Froude number after the V = V3 Ac 3 = V3 by 3 = (2 m/s)(1 m)(3 m) = 6 m 3 /s
y1 =
V3 2 m/s y3 = (3 m ) = 4.8 m V1 1.25 m/s
Fr3 =
V3 gy 3
=
2 m/s (9.81 m/s 2 )(3 m)
= 0.3687
The flow dept, velocity, and Froude number before the jump are
)
(
(
)
y2 = 0.5 y3 −1 + 1 + 8Fr32 = 0.5(3 m) −1 + 1 + 8 × 0.3687 2 = 0.6671 m ≅ 0.667 m V2 =
y3 3m V3 = (2 m/s ) = 8.994 m/s y2 0.6671 m
Fr2 =
V2 gy 2
=
8.994 m/s (9.81 m/s 2 )(0.6671 m)
= 3.516
which is greater than 1, and thus the flow before the jump is indeed supercritical. The head loss and the fraction of mechanical energy dissipated during hydraulic jump are hL = y 2 − y 3 +
V 22 − V32 (8.994 m/s) 2 - (2 m/s) 2 = (0.6671 m) - (3 m) + = 1.586 m 2g 2(9.81 m/s 2 )
hL hL 1.586 m = = = 0.331 2 E s 2 y 2 (1 + Fr2 / 2) (0.6671 m)(1 + 3.516 2 / 2) Note that this is a fairly “strong” hydraulic jump, wasting 33.1% of the mechanical energy of the fluid.
Dissipation ratio =
Discussion
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Chapter 13 Open-Channel Flow 13-138 Solution Water from a lake is discharged through a sluice gate into a channel where uniform flow conditions are established, and then undergoes a hydraulic jump. The flow depth, velocity, and Froude number after the jump are to be determined. Assumptions 1 The flow is steady. 2 The channel is sufficiently wide so that the end effects are negligible. 3 The effects of channel slope on hydraulic jump are negligible.
Sluice gate
Hydraulic jump
y1 = 5 m
Properties The Manning coefficient for an open channel made of finished concrete is n = 0.012 (Table 131).
a =0.5 m
Analysis For free outflow, we only need the depth ratio y1/a to determine the discharge coefficient,
y3 y2
y1 5m = = 10 a 0.5 m
The corresponding discharge coefficient is determined from Fig. 13-38 to be Cd = 0.58. Then the discharge rate per m width (b = 1 m) becomes V = C d ba 2 gy1 = 0.58 (1 m)(0.5 m) 2 (9.81 m/s 2 )(5 m) = 2.872 m 3 /s
For wide channels, hydraulic radius is the flow depth and thus Rh = y2. Then the flow depth in uniform flow after the gate is determined from the Manning’s equation to be a 1 m 1/3 / s V = Ac R h2 / 3 S 01 / 2 → 2.872 m 3 /s = [(1 m) y 2 ]( y 2 ) 2/3 0.004 1 / 2 n 0.012 It gives y2 = 0.6948 m, which is also the flow depth before water undergoes a hydraulic jump. The flow velocity and Froude number in uniform flow are V 2.872 m 3 /s V2 = = = 4.134 m/s b y 2 (1 m) ( 0.6948 m) Fr2 =
V2 gy 2
=
4.134 m/s (9.81 m/s 2 )(0.6948 m)
= 1.584
Then the flow depth, velocity, and Froude number after the jump (state 3) become y 3 = 0.5 y 2 ⎛⎜ − 1 + 1 + 8Fr22 ⎞⎟ = 0.5(0.6948 m)⎛⎜ − 1 + 1 + 8 × 1.584 2 ⎞⎟ = 1.25 m ⎝ ⎠ ⎝ ⎠ V3 =
y2 0.6948 m V2 = (4.134 m/s ) = 2.30 m/s y3 1.25 m
Fr3 =
V3 gy 23
=
2.30 m/s (9.81 m/s 2 )(1.25 m)
= 0.659
Discussion This is a relatively “mild” jump. It can be shown that the head loss during hydraulic jump is 0.049 m, which corresponds to an energy dissipation ratio of 3.1%.
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Chapter 13 Open-Channel Flow 13-139 Solution Water is discharged from a dam into a wide spillway to reduce the risk of flooding by dissipating a large fraction of mechanical energy via hydraulic jump. For specified flow depths, the velocities before and after the jump, and the mechanical power dissipated per meter with of the spillway are to be determined. Assumptions negligible.
1 The flow is steady or quasi-steady. 2 The channel is sufficiently wide so that the end effects are
Properties
The density of water is 1000 kg/m3.
Analysis The Froude number and velocity before the jump are y2 4m = 0.5⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ → = 0.5⎛⎜ − 1 + 1 + 8Fr12 ⎞⎟ ⎝ ⎠ ⎝ ⎠ y1 0.5 m which gives Fr1 = 6. Also, from the definition of Froude number,
hL
V1 = Fr1 gy1 = (6) (9.81 m/s 2 )(0.5 m) = 13.3 m/s
Velocity and Froude number after the jump are
V
V2 FIGURE 13-51 y2 = 4 m Control volume that encloses the y1 = 0.5 m hydraulic jump.
y 0.5 m (13.3 m/s ) = 1.66 m/s V 2 = 1 V1 = y2 4m Fr2 =
V2 gy 2
=
1.66 m/s (9.81 m/s 2 )(4 m)
Energy line
1
(1)
(2)
= 0.265
The head loss is determined from the energy equation to be h L = y1 − y 2 +
V12 − V 22 (13.3 m/s) 2 - (1.66 m/s) 2 = (0.5 m) - (4 m) + = 5.36 m 2g 2(9.81 m/s 2 )
The volume and mass flow rates of water per m width are V = V1 Ac1 = V1by1 = (13.3 m/s )(1 m)(0.5 m) = 6.64 m 3 /s m = ρV = (1000 kg/m 3 )(6.64 m 3 /s) = 6640 kg/s
Then the dissipated mechanical power becomes ⎛ 1 kN E dissipated = m gh L = (6640 kg/s)(9.81 m/s 2 )(5.36 m)⎜⎜ 2 ⎝ 1000 kg ⋅ m/s
Discussion case.
⎞ ⎟ = 349 kNm/s = 349 kW ⎟ ⎠
The results show that the hydraulic jump is a highly dissipative process, wasting 349 kW of power in this
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Chapter 13 Open-Channel Flow 13-140 Solution The flow rate in an open channel is to be measured using a sharp-crested rectangular weir. For a measured value of flow depth upstream, the flow rate is to be determined.
H = 0.60 m
V1 y1
Pw = 1.1 m
Sharp-crested rectangular weir
Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis
The weir head is given to be H = 0.60 m. The discharge coefficient of the weir is
C wd ,rec = 0.598 + 0.0897
H 0.60 m = 0.598 + 0.0897 = 0.6469 Pw 1.1 m
The condition H/Pw < 2 is satisfied since 0.60/1.1 = 0.55. Then the water flow rate through the channel becomes 2 V = C wd ,rec b 2 g H 3 / 2 3 2 = (0.6469) (6 m) 2(9.81 m/s 2 ) (0.60 m ) 3 / 2 3 3 = 5.33 m /s
Discussion The upstream velocity and the upstream velocity head are V12 (0.522 m/s) 2 V 5.33 m 3 /s V1 = = = 0.522 m/s and = = 0.014 m 2 g 2(9.81 m/s 2 ) by1 (6 m)(1.70 m) This is 2.3% of the weir head, which is negligible. When the upstream velocity head is considered, the flow rate becomes 5.50 m3/s, which is about 3 percent higher than the value determined above. Therefore, the assumption of negligible velocity head is reasonable in this case.
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Chapter 13 Open-Channel Flow 13-141E Solution The flow rates in two open channels are to be measured using a sharp-crested weir in one and a broadcrested rectangular weir in the other. For identical flow depths, the flow rates through both channels are to be determined.
Sharp-crested rectangular weir
Discharge
V1
y1 = 5 ft
y1 = 5 ft Pw = 2 ft
Pw = 2 ft Broad-crested weir
Assumptions 1 The flow is steady. 2 The upstream velocity head is negligible. 3 The channel is sufficiently wide so that the end effects are negligible. Analysis
The weir head is
H = y1 − Pw = 5.0 ft − 2.0 ft = 3.0 ft
The condition H/Pw < 2 is satisfied since 3.0/2.0 = 1.5. The discharge coefficients of the weirs are Sharp-crested weir: C wd ,sharp = 0.598 + 0.0897
H 3.0 ft = 0.598 + 0.0897 = 0.7326 Pw 2.0 ft
2 2 Vsharp = C wd ,sharp b 2 g H 3 / 2 = (0.7326) (12 ft) 2(32.2 ft/s 2 ) (3.0 ft ) 3 / 2 = 244 m 3 /s 3 3
Broad-crested weir:
C wd ,broad =
0.65 1 + H / Pw
=
⎛2⎞ Vbroad = C wd ,broad b g ⎜ ⎟ ⎝ 3⎠
0.65 1 + (3.0 ft)/(2.0 ft) 3/ 2
= 0.4111
H 3 / 2 = (0.4111)(12 ft)(2/3) 3/2 32.2 ft/s 2 (3.0 ft ) 3 / 2 = 79.2 m 3 /s
Discussion Note that the flow rate in the channel with the broad-crested weir is much less than the channel with the sharp-crested weir. Also, if the upstream velocity is taken into consideration, the flow rate would be 270 ft3/s (11% difference) for the channel with the sharp-crested weir, and 80.3 ft3/s (1% difference) for the one with broad-crested weir. Therefore, the assumption of negligible dynamic head is not quite appropriate for the channel with the sharp-crested weir.
Design and Essay Problems
13-142 to 13-143 Solution Students’ essays and designs should be unique and will differ from each other.
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Chapter 14 Turbomachinery
Solutions Manual for Fluid Mechanics: Fundamentals and Applications by Çengel & Cimbala
CHAPTER 14 TURBOMACHINERY
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Chapter 14 Turbomachinery General Problems
14-1C Solution
We are to discuss energy producing and energy absorbing devices.
Analysis A more common term for an energy producing turbomachine is a turbine. Turbines extract energy from the moving fluid, and convert that energy into useful mechanical energy in the surroundings, usually in the form of a rotating shaft. Thus, the phrase “energy producing” is from a frame of reference of the fluid – the fluid loses energy as it drives the turbine, producing energy to the surroundings. On the other hand, a more common term for an energy absorbing turbomachine is a pump. Pumps absorb mechanical energy from the surroundings, usually in the form of a rotating shaft, and increase the energy of the moving fluid. Thus, the phrase “energy absorbing” is from a frame of reference of the fluid – the fluid gains or absorbs energy as it flows through the pump. Discussion From the frame of reference of the surroundings, a pump absorbs energy from the surroundings, while a turbine produces energy to the surroundings. Thus, you may argue that the terminology also holds for the frame of reference of the surroundings. This alternative explanation is also acceptable.
14-2C Solution
We are to discuss the differences between fans, blowers, and compressors.
Analysis A fan is a gas pump with relatively low pressure rise and high flow rate. A blower is a gas pump with relatively moderate to high pressure rise and moderate to high flow rate. A compressor is a gas pump designed to deliver a very high pressure rise, typically at low to moderate flow rates. Discussion
The boundaries between these three types of pump are not always clearly defined.
14-3C Solution
We are to list examples of fans, blowers, and compressors.
Analysis Common examples of fans are window fans, ceiling fans, fans in computers and other electronics equipment, radiator fans in cars, etc. Common examples of blowers are leaf blowers, hair dryers, air blowers in furnaces and automobile ventilation systems. Common examples of compressors are tire pumps, refrigerator and air conditioner compressors. Discussion
Students should come up with a diverse variety of examples.
14-4C Solution We are to discuss the difference between a positive-displacement turbomachine and a dynamic turbomachine. Analysis A positive-displacement turbomachine is a device that contains a closed volume; energy is transferred to the fluid (pump) or from the fluid (turbine) via movement of the boundaries of the closed volume. On the other hand, a dynamic turbomachine has no closed volume; instead, energy is transferred to the fluid (pump) or from the fluid (turbine) via rotating blades. Examples of positive-displacement pumps include well pumps, hearts, some aquarium pumps, and pumps designed to release precise volumes of medicine. Examples of positive-displacement turbines include water meters and gas meters in the home. Examples of dynamic pumps include fans, centrifugal blowers, airplane propellers, centrifugal water pumps (like in a car engine), etc. Examples of dynamic turbines include windmills, wind turbines, turbine flow meters, etc. Discussion
Students should come up with a diverse variety of examples.
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Chapter 14 Turbomachinery 14-5C Solution efficiency.
We are to discuss the difference between brake horsepower and water horsepower, and then discuss pump
Analysis In turbomachinery terminology, brake horsepower is the power actually delivered to the pump through the shaft. (One may also call it “shaft power”.) On the other hand, water horsepower is the useful portion of the brake horsepower that is actually delivered to the fluid. Water horsepower is always less than brake horsepower; hence pump efficiency is defined as the ratio of water horsepower to brake horsepower. Discussion horsepower.
14-6C Solution efficiency.
For a turbine, efficiency is defined in the opposite way, since brake horsepower is less than water
We are to discuss the difference between brake horsepower and water horsepower, and then discuss turbine
Analysis In turbomachinery terminology, brake horsepower is the power actually delivered by the turbine to the shaft. (One may also call it “shaft power”.) On the other hand, water horsepower is the power extracted from the water flowing through the turbine. Water horsepower is always greater than brake horsepower; because of inefficiencies; hence turbine efficiency is defined as the ratio of brake horsepower to water horsepower. Discussion horsepower.
14-7C Solution
For a pump, efficiency is defined in the opposite way, since brake horsepower is greater than water
We are to explain the “extra” term in the Bernoulli equation in a rotating reference frame.
Analysis A rotating reference frame is not an inertial reference frame. When we move outward in the radial direction, the absolute velocity at this location is faster due to the rotating body, since vθ is equal to ωr. When solving a turbomachinery problem in a rotating reference frame, we use the relative fluid velocity (velocity relative to the rotating reference frame). Thus, in order for the Bernoulli equation to be physically correct, we must subtract the absolute velocity of the rotating body so that the equation applies to an inertial reference frame. This accounts for the “extra” term. Discussion The Bernoulli equation is the same physical equation in either the absolute or the rotating reference frame, but it is more convenient to use the form with the extra term in turbomachinery applications.
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Chapter 14 Turbomachinery 14-8 Solution water pump.
We are to determine how the average speed at the outlet compares to the average speed at the inlet of a
Assumptions
1 The flow is steady (in the mean). 2 The water is incompressible.
Analysis
Conservation of mass requires that the mass flow rate in equals the mass flow rate out. Thus,
Conservation of mass:
m in = ρinVin Ain = m out = ρ outVout Aout
or, since the cross-sectional area is proportional to the square of diameter, 2
Vout
⎛ D ⎞ ρ ⎛D ⎞ = Vin in ⎜ in ⎟ = Vin ⎜ in ⎟ ρout ⎝ Dout ⎠ ⎝ Dout ⎠
2
(1)
(a) For the case where Dout < Din, Vout must be greater than Vin. (b) For the case where Dout = Din, Vout must be equal to Vin. (c) For the case where Dout > Din, Vout must be less than Vin. Discussion A pump does not necessarily increase the speed of the fluid passing through it. In fact, the average speed through the pump can actually decrease, as it does here in part (c).
14-9 Solution For an air compressor with equal inlet and outlet areas, and with both density and pressure increasing, we are to determine how the average speed at the outlet compares to the average speed at the inlet. Assumptions
1 The flow is steady.
Analysis Conservation of mass requires that the mass flow rate in equals the mass flow rate out. The cross-sectional areas of the inlet and outlet are the same. Thus, Conservation of mass:
m in = ρinVin Ain = m out = ρ outVout Aout
or Vout = Vin
ρin ρout
(1)
Since ρin < ρout, Vout must be less than Vin. Discussion A compressor does not necessarily increase the speed of the fluid passing through it. In fact, the average speed through the pump can actually decrease, as it does here.
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Chapter 14 Turbomachinery Pumps
14-10C Solution
We are to list and define the three categories of dynamic pumps.
Analysis The three categories are: Centrifugal flow pump – fluid enters axially (in the same direction as the axis of the rotating shaft) in the center of the pump, but is discharged radially (or tangentially) along the outer radius of the pump casing. Axial-flow pump – fluid enters and leaves axially, typically only along the outer portion of the pump because of blockage by the shaft, motor, hub, etc. Mixed-flow pump – intermediate between centrifugal and axial, with the flow entering axially, not necessarily in the center, but leaving at some angle between radially and axially. Discussion There are also some non-rotary dynamic pumps, such as jet pumps and electromagnetic pumps, that are not discussed in this text.
14-11C Solution (a) False: Actually, backward-inclined blades yield the highest efficiency. (b) True: The pressure rise is higher, but at the cost of less efficiency. (c) True: In fact, this is the primary reason for choosing forward-inclined blades. (d) False: Actually, the opposite is true – a pump with forward-inclined blades usually has more blades, but they are usually smaller.
14-12C Solution
We are to choose which pump location is better and explain why.
Analysis The two systems are identical except for the location of the pump (and some minor differences in pipe layout). The overall length of pipe, number of elbows, elevation difference between the two reservoir free surfaces, etc. are the same. Option (a) is better because it has the pump at a lower elevation, increasing the net positive suction head, and lowering the possibility of pump cavitation. In addition, the length of pipe from the lower reservoir to the pump inlet is smaller in Option (a), and there is one less elbow between the lower reservoir and the pump inlet, thereby decreasing the head loss upstream of the pump – both of which also increase NPSH, and reduce the likelihood of cavitation. Discussion Another point is that if the pump is not self-priming, Option (b) may run into start-up problems if the free surface of the lower reservoir falls below the elevation of the pump inlet. Since the pump in Option (a) is below the reservoir, self-priming is not an issue.
14-13C Solution
We are to define and discuss NPSH and NPSHrequired.
Analysis Net positive suction head (NPSH) is defined as the difference between the pump’s inlet stagnation pressure head and the vapor pressure head, ⎛ P V2 ⎞ P NPSH = ⎜ + − v ⎟ ρ g g ρ g 2 ⎝ ⎠ pump inlet
We may think of NPSH as the actual or available net positive suction head. On the other hand, required net positive suction head (NPSHrequired) is defined as the minimum NPSH necessary to avoid cavitation in the pump. As long as the actual NPSH is greater than NPSHrequired, there should be no cavitation in the pump. Discussion Although NPSH and NPSHrequired are measured at the pump inlet, cavitation (if present) happens somewhere inside the pump, typically on the suction surface of the rotating pump impeller blades.
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Chapter 14 Turbomachinery 14-14C Solution (a) True: As volume flow rate increases, not only does NPSHrequired increase, but the available NPSH decreases, increasing the likelihood that NPSH will drop below NPSHrequired and cause cavitation to occur. (b) False: NPSHrequired is not a function of water temperature, although available NPSH does depend on water temperature. (c) False: Available NPSH actually decreases with increasing water temperature, making cavitation more likely to occur. (d) False: Actually, warmer water causes cavitation to be more likely. The best way to think about this is that warmer water is already closer to its boiling point, so cavitation is more likely to happen in warm water than in cold water.
14-15C Solution
We are to explain why dissimilar pumps should not be arranged in series or in parallel.
Analysis Arranging dissimilar pumps in series can create problems because the volume flow rate through each pump must be the same, but the overall pressure rise is equal to the pressure rise of one pump plus that of the other. If the pumps have widely different performance curves, the smaller pump may be forced to operate beyond its free delivery flow rate, whereupon it acts like a head loss, reducing the total volume flow rate. Arranging dissimilar pumps in parallel can create problems because the overall pressure rise must be the same, but the net volume flow rate is the sum of that through each branch. If the pumps are not sized properly, the smaller pump may not be able to handle the large head imposed on it, and the flow in its branch could actually be reversed; this would inadvertently reduce the overall pressure rise. In either case, the power supplied to the smaller pump would be wasted. Discussion
If the pumps are not significantly dissimilar, a series or parallel arrangement of the pumps might be wise.
14-16C Solution (a) True: The maximum volume flow rate occurs when the net head is zero, and this “free delivery” flow rate is typically much higher than that at the BEP. (b) True: By definition, there is no flow rate at the shutoff head. Thus the pump is not doing any useful work, and the efficiency must be zero. (c) False: Actually, the net head is typically greatest near the shutoff head, at zero volume flow rate, not near the BEP. (d) True: By definition, there is no head at the pump’s free delivery. Thus, the pump is working against no “resistance”, and is therefore not doing any useful work, and the efficiency must be zero.
14-17C Solution
We are to discuss ways to improve the cavitation performance of a pump, based on the equation for NPSH.
Analysis
NPSH is defined as
⎛ P V2 ⎞ P + − v NPSH = ⎜ ⎟ ⎝ ρ g 2 g ⎠ pump inlet ρ g
(1)
To avoid cavitation, NPSH must be increased as much as possible. For a given liquid at a given temperature, the vapor pressure head (last term on the right side of Eq. 1) is constant. Hence, the only way to increase NPSH is to increase the stagnation pressure head at the pump inlet. We list several ways to increase the available NPSH: (1) Lower the pump or raise the inlet reservoir level. (2) Use a larger diameter pipe upstream of the pump. (3) Re-route the piping system such that fewer minor losses (elbows, valves, etc.) are encountered upstream of the pump. (4) Shorten the length of pipe upstream of the pump. (5) Use a smoother pipe. (6) Use elbows, valves, inlets, etc. that have smaller minor loss coefficients. Suggestion (1) raises NPSH by increasing the hydrostatic component of pressure at the pump inlet. Suggestions (2) through (6) raise NPSH by lowering the irreversible head losses, thereby increasing the pressure at the pump inlet. Discussion By definition, when the available NPSH falls below the required NPSH, the pump is prone to cavitation, which should be avoided if at all possible.
14-6 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-18C Solution (a) False: Since the pumps are in series, the volume flow rate through each pump must be the same: V = V1 = V2 . (b) True: The net head increases by H1 through the first pump, and then by H2 through the second pump. The overall rise in net head is thus the sum of the two. (c) True: Since the pumps are in parallel, the total volume flow rate is the sum of the individual volume flow rates. (d) False: For pumps in parallel, the change in pressure from the upstream junction to the downstream junction is the same regardless of which parallel branch is under consideration. Thus, even though the volume flow rate may not be the same in each branch, the net head must be the same: H = H1 = H2.
14-19C Solution
Shutoff head Pump performance curve
We are to label several items on the provided plot.
Analysis The figure is re-drawn here, and the requested items are labeled.
Havailable H System curve
Discussion Also labeled are the available net head, corresponding to the pump performance curve, and the required net head, corresponding to the system curve. The intersection of these two curves is the operating point of the pump.
14-20 Solution equation.
Operating point
Free delivery
Hrequired 0
V
0
We are to determine which free surface is at higher elevation, and justify our answer with the energy
Analysis It is simplest to consider zero-flow conditions ( V = 0), at which we see that the required net head is positive. This implies that, even when there is no flow between the two tanks, the pump would need to provide some net head just to overcome the pressure differences. Since there is no flow, pressure differences can come only from gravity. Hence, the outlet tank’s free surface must be higher than that of the inlet tank. Mathematically, we apply the energy equation in head form between the inlet tank’s free surface (1) and the outlet tank’s free surface (2),
Energy equation at zero flow conditions: H required = hpump,u =
2 2 P2 − P1 α 2 V2 − α1 V1 + ( z2 − z1 ) + hturbine + hL ,total + ρg 2g
(1)
Since both free surfaces are at atmospheric pressure, P1 = P2 = Patm, and the first term on the right side of Eq. 1 vanishes. Furthermore, since there is no flow, V1 = V2 = 0, and the second term vanishes. There is no turbine in the control volume, so the second-to-last term is zero. Finally, there are no irreversible head losses since there is no flow, and the last term is also zero. Equation 1 reduces to H required = hpump = ( z2 − z1 )
(2)
Since Hrequired is positive on Fig. P14-19 at V = 0, the quantity (z2 – z1) must also be positive by Eq. 2. Thus we have shown mathematically that the outlet tank’s free surface is higher in elevation than that of the inlet tank. Discussion If the reverse were true (outlet tank free surface lower than inlet tank free surface), Hrequired at V = 0 would be negative, implying that the pump would need to supply enough negative net head to hold back the natural tendency of the water to flow from higher to lower elevation. In reality, the pump would not be able to do this unless it were spun backwards.
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Chapter 14 Turbomachinery 14-21 Solution We are to discuss what would happen to the pump performance curve, the system curve, and the operating point if the free surface of the outlet tank were raised to a higher elevation. Analysis The pump is the same pump regardless of the locations of the inlet and outlet tanks’ free surfaces; thus, the pump performance curve does not change. The energy equation is
H required = hpump,u =
P2 − P1 α 2V2 2 − α1V12 + + ( z2 − z1 ) + hturbine + hL ,total ρg 2g
Since the only thing that changes is the elevation difference, Eq. 1 shows that Hrequired shifts up as (z2 – z1) increases. Thus, the system curve rises linearly with elevation increase. A plot of H versus V is plotted, and the new operating point is labeled. Because of the upward shift of the system curve, the operating point moves to a lower value of volume flow rate.
(1)
Pump performance curve (does not change) Havailable
New operating point
H Original operating point
New system curve
Free delivery
Discussion The shift of operating point to lower V agrees with our physical intuition. Namely, as we raise the elevation of the outlet, the pump has to do more work to overcome gravity, and we expect the flow rate to decrease accordingly.
Hrequired Original system curve 0
V
0
14-22 Solution We are to discuss what would happen to the pump performance curve, the system curve, and the operating point if a valve changes from 100% to 50% open. Analysis The pump is the same pump regardless of the locations of the inlet and outlet tanks’ free surfaces; thus, the pump performance curve does not change. The energy equation is H required = hpump,u =
P2 − P1 α 2V2 2 − α1V12 + + ( z2 − z1 ) + hturbine + hL ,total ρg 2g
Since both free surfaces are open to the atmosphere, the pressure term vanishes. Since both V1 and V2 are negligibly small at the free surface (the tanks are large), the second term on the right also vanishes. The elevation difference (z2 – z1) does not change, and so the only term in Eq. 1 that is changed by closing the valve is the irreversible head loss term. We know that the minor loss associated with a valve increases significantly as the valve is closed. Thus, the system curve (the curve of Hrequired versus V ) increases more rapidly with volume flow rate (has a larger slope) when the valve is partially closed. A sketch of H versus V is plotted, and the new operating point is labeled. Because of the higher system curve, the operating point moves to a lower value of volume flow rate, as indicated on the figure. I.e., the volume flow rate decreases.
(1)
Pump performance curve (does not change) New operating point Havailable H New system curve
Original operating point
Free delivery Hrequired 0
0
Original system curve
V
Discussion The shift of operating point to lower V agrees with our physical intuition. Namely, as we close the valve somewhat, the pump has to do more work to overcome the losses, and we expect the flow rate to decrease accordingly.
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Chapter 14 Turbomachinery 14-23 Solution We are to create a qualitative plot of pump net head versus pump capacity.
Shutoff head Pump performance curve
Analysis The result is shown in the figure, and the requested items are labeled. Also labeled are the available net head, corresponding to the pump performance curve, and the required net head, corresponding to the system curve. The intersection of these two curves is the operating point of the pump. Note that since the elevation of the outlet is lower than that of the free surface of the inlet tank, the required net head must be negative at zero flow rate conditions, as sketched, implying that the pump holds back the natural tendency of the water to flow from higher to lower elevation. Only at higher flow rates does the system curve rise to positive values of Hrequired. Discussion A real pump cannot produce negative net head at zero volume flow rate unless its blades are spun in the opposite direction than that for which they are designed.
Operating point
Havailable
H
Free delivery
Hrequired 0
V
0 System curve
14-9 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-24 Solution
We are to estimate the volume flow rate through a piping system.
Assumptions 1 Since the reservoir is large, the flow is nearly steady. 2 The water is incompressible. 3 The water is at room temperature. 4 The flow in the pipe is fully developed and turbulent, with α = 1.05.
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Properties
Analysis By definition, at free delivery conditions, the net head across the pump is zero. Thus, there is no loss or gain of pressure across the pump, and we can essentially ignore it in the calculations here. We apply the head form of the steady energy equation from location 1 to location 2, H required = hpump = 0 =
2 2 P2 − P1 α 2V2 − V1 + + ( z2 − z1 ) + hturbine + hL ,total ρg 2g
(1)
where the pressure term vanishes since the free surface at location 1 and at the exit (location 2) are both open to the atmosphere. The inlet velocity term disappears since V1 is negligibly small at the free surface. Thus, Eq. 1 reduces to a balance between supplied potential energy head (z1 – z2), kinetic energy head at the exit α2V22/2g, and irreversible head losses,
( z1 − z2 ) =
α 2V2 2 2g
+ hL ,total
(2)
The total irreversible head loss in Eq. 2 consists of both major and minor losses. We split the minor losses into those associated with the mean velocity V through the pipe, and the minor loss associated with the contraction, based on exit velocity V2,
( z1 − z2 ) = where
∑K
L
α 2V2 2 2g
+
⎞ V2 2 V2 ⎛ L f + K ⎜ ∑ L ⎟ + K L,contraction 2 g ⎝ D pipe ⎠ 2 g
(3)
= 0.50 + 2(2.4) + 3(0.90) = 8.0, and K L ,contraction = 0.15.
pipe
By conservation of mass, VA = V2 A2
→
⎛ D⎞ A =V ⎜ V2 = V ⎟ A2 ⎝ D2 ⎠
2
(4)
Substitution of Eq. 4 into Eq. 3 yields 4 ⎞ ⎛ D⎞ V2 ⎛ L ⎜ ( z1 − z2 ) = ⎜ f + ∑ K L + ⎜ ⎟ (α 2 + K L,contraction ) ⎟⎟ 2g D pipe ⎝ D2 ⎠ ⎝ ⎠
(5)
Equation 5 is an implicit equation for V since the Darcy friction factor is a function of Reynolds number Re = ρVD/μ, as obtained from either the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 1.911 m/s, or to three significant digits, V = 1.91 m/s, from which the volume flow rate is
π ( 0.020 m ) π D2 V = V = (1.911 m/s ) = 6.01× 10 −4 m 3 /s 4 4 2
(6)
In more common units, V = 36.0 Lpm (liters per minute). The Reynolds number is 3.81 × 104. Discussion Since there is no net head across the pump at free delivery conditions, the pump could be removed (inlet and outlet pipes connected together without the pump), and the flow rate would be the same. Another way to think about this is that the pump’s efficiency is zero at the free delivery operating point, so it is doing no useful work.
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Chapter 14 Turbomachinery 14-25 Solution
We are to calculate the volume flow rate through a piping system in which the pipe is rough.
Assumptions 1 Since the reservoir is large, the flow is nearly steady. 2 The water is incompressible. 3 The water is at room temperature. 4 The flow in the pipe is fully developed and turbulent, with α = 1.05. Properties
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Analysis The relative pipe roughness is ε/D = (0.050 cm)/(2.0 cm) = 0.025 (very rough, as seen on the Moody chart). The calculations are identical to that of the previous problem, except for the pipe roughness. The result is V = 1.597 m/s, or to three significant digits, V = 1.60 m/s, from which the volume flow rate is 5.02 × 10-4 m3/s, or V = 30.1 Lpm. The Reynolds number is 3.18 × 104. The volume flow rate is lower by about 16%. This agrees with our intuition, since pipe roughness leads to more pressure drop at a given flow rate. Discussion calculations.
If the calculations of the previous problem are done on a computer, it is trivial to change ε for the present
14-26
For a given pump and piping system, we are to calculate the volume flow rate and compare with that Solution calculated for Problem 14-24. Assumptions 1 Since the reservoir is large, the flow is nearly steady. 2 The water is incompressible. 3 The water is at room temperature. 4 The flow in the pipe is fully developed and turbulent, with α = 1.05. Properties
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Analysis The calculations are identical to those of Problem 14-24 except that the pump’s net head is not zero as in Problem 14-24, but instead is given in the problem statement. At the operating point, we match Havailable to Hrequired, yielding H available = H required
→
4 ⎞ ⎛ D ⎞ V2 ⎛ L 2 ⎜ f + ∑ KL + ⎜ H 0 − aV = ⎟ (α 2 + K L ,contraction ) ⎟ − ( z1 − z2 ) ⎟ 2 g ⎜ D pipe ⎝ D2 ⎠ ⎝ ⎠
(1)
We re-write the second term on the left side of Eq. 1 in terms of average pipe velocity V instead of volume flow rate, since V = VπD2/4, and solve for V, V=
H 0 + ( z1 − z2 ) 4 ⎞ ⎛ D⎞ 1 ⎛ L π 2 D4 ⎜f + ∑ KL + ⎜ ⎟ (α 2 + K L ,contraction ) ⎟ + a ⎟ 2 g ⎜ D pipe 16 ⎝ D2 ⎠ ⎝ ⎠
(2)
Equation 2 is an implicit equation for V since the Darcy friction factor is a function of Reynolds number Re = ρVD/μ, as obtained from either the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 2.846 m/s, from which the volume flow rate is
π ( 0.020 m ) π D2 m3 V = V = ( 2.846 m/s ) = 8.942 × 10−4 4 4 s 2
(3)
In more common units, V = 53.6 Lpm. This represents an increase of about 49% compared to the flow rate of Problem 1424. This agrees with our expectations – adding a pump in the line produces a higher flow rate. Discussion Although there was a pump in Problem 14-24 as well, it was operating at free delivery conditions, implying that it was not contributing anything to the flow – that pump could be removed from the system with no change in flow rate. Here, however, the net head across the pump is about 5.34 m, implying that it is contributing useful head to the flow (in addition to the gravity head already present).
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Chapter 14 Turbomachinery 14-27E Solution We are to calculate pump efficiency and estimate the BEP conditions. Properties
The density of water at 77oF is 62.24 lbm/ft3.
Analysis
(a) Pump efficiency is
η pump =
Pump efficiency:
ρ gV H
(1)
bhp
We show the second row of data (at V = 4.0 gpm) as an example – the rest are calculated in a spreadsheet for convenience, gal ⎞ ( 62.24 lbm/ft ) ⎛⎜⎝ 32.2 sft ⎞⎟⎠ ⎛⎜⎝ 4.0 min ⎟ (18.5 ft ) ⎛ 0.1337 ft ⎠ 3
η pump =
2
0.064 hp
⎜ ⎝
3
gal
Pump performance data for water at 77oF. V (gpm) 0.0 4.0 8.0 12.0 16.0 20.0 24.0
H (ft) 19.0 18.5 17.0 14.5 10.5 6.0 0
ηpump
bhp (hp) 0.06 0.064 0.069 0.074 0.079 0.08 0.078
(%) 0.0 29.2 49.7 59.3 53.6 37.8 0.0
⎞ ⎛ 1 min ⎞ ⎛ lbf ⋅ s 2 ⎞ ⎛ hp ⋅ s ⎞ ⎟⎜ ⎟⎜ ⎟×⎜ ⎟ = 0.292 ⎠ ⎝ 60 s ⎠ ⎝ 32.2 lbm ⋅ ft ⎠ ⎝ 550 ft ⋅ lbf ⎠
or 29.2%. The results for all rows are shown in the table. (b) The best efficiency point (BEP) occurs at approximately the fourth row of data: V * = 12.0 gpm, H* = 14.5 ft of head, bhp* = 0.074 hp, and ηpump* = 59.3%. Discussion
A more precise BEP could be obtained by curve-fitting the data, as in Problem 14-29.
14-28 Solution We are to convert the pump performance data to metric units and calculate pump efficiency. Properties
The density of water at T = 20oC is 998.0 kg/m3.
Analysis The conversions are straightforward, and the results are shown in the table. A sample calculation of the pump efficiency for the second row of data is shown below:
V (Lpm) 0.0 15.1 30.3 45.4 60.6 75.7 90.9
H (m) 5.79 5.64 5.18 4.42 3.20 1.83 0.00
bhp (W) 44.7 47.7 51.5 55.2 58.9 59.7 58.2
ηpump (%) 0.0 29.2 49.7 59.3 53.6 37.8 0.0
( 998.0 kg/m )( 9.81 m/s ) (15.1 L/min )( 5.64 m ) ⎛ 3
η pump =
Pump performance data for water at 77oF.
2
47.7 W
1 m3 ⎞ ⎛ 1 min ⎞ ⎛ N ⋅ s 2 ⎞ ⎛ W ⋅ s ⎞ ⎜ ⎟⎜ ⎟⎜ ⎟×⎜ ⎟ = 0.292 = 29.2% ⎝ 1000 L ⎠ ⎝ 60 s ⎠ ⎝ kg ⋅ m ⎠ ⎝ N ⋅ m ⎠
The pump efficiency data are identical to those of the previous problem, as they must be, regardless of the system of units. Discussion If the calculations of the previous problem are done on a computer, it is trivial to convert to metric units in the present calculations.
14-12 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-29E [Also solved using EES on enclosed DVD] Solution We are to generate least-squares polynomial curve fits of a pump’s performance curves, plot the curves, and calculate the BEP.
The density of water at 77oF is 62.24 lbm/ft3.
20
Analysis The efficiencies for each data point in Table P14-27 are calculated in Problem 14-27. We use regression analysis to generate the least-squares fits. The equation and coefficients for H are
15
Properties
H = H 0 − aV 2
H 0 = 19.0774 ft
a = 0.032996 ft/gpm2
Or, to 3 digits of precision, H 0 = 19.1 ft
a1 = 0.00175 hp/gpm
0.08
10
bhp (hp)
5
bhp0 = 0.0587 hp a2 = -3.72 ×10-5 hp/gpm2
a1 = 8.21 %/gpm
0.04 0
5
a3 = -0.00546 %/gpm3
The tabulated data are plotted in Fig. 1 as symbols only. The fitted data are plotted on the same plots as lines only. The agreement is excellent. The best efficiency point is obtained by differentiating the curvefit expression for ηpump with respect to volume flow rate, and setting the derivative to zero (solving the resulting quadratic equation for V * ), dη pump = a1 + 2a2V + 3a3V 2 =0 dV
→
V * = 12.966 GPM ≈ 13.0 gpm
At this volume flow rate, the curve-fitted expressions for H, bhp, and ηpump yield the operating conditions at the best efficiency point (to three digits each): V * = 13.0 gpm,
H * = 13.5 ft,
10
V
ηpump,0 = 0.0523%
a2 = -0.210 %/gpm2
bhp* = 0.0752 hp,
η * = 59.2%
Discussion This BEP is more precise than that of Problem 14-27 because of the curve fit. The other root of the quadratic is negative − obviously not the correct choice.
0.06
0
The equation and coefficients for ηpump are
ηpump = ηpump,0 + a1V + a2V 2 + a3V 3
0.10
H (ft)
a = 0.0330 ft/gpm2
The equation and coefficients for bhp are bhp = bhp0 + a1V + a2V 2
0.12
15
20
25
(gpm) (a)
60 50
ηpump
40
(%) 30 20 10 0 0
5
10
V
15
20
25
(gpm)
(b) FIGURE 1 Pump performance curves: (a) H and bhp versus V , and (b) ηpump versus V .
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Chapter 14 Turbomachinery
14-30E Solution
For a given pump and system requirement, we are to estimate the operating point.
Assumptions
1 The flow is steady. 2 The water is at 77oF and is incompressible.
Analysis The operating point is the volume flow rate at which Hrequired = Havailable. We set the given expression for Hrequired to the curve fit expression of Problem 14-29, Havailable = H0 − a V 2 , and obtain H 0 − ( z2 − z1 )
V =
Operating point:
a+b
=
19.0774 ft − 15.5 ft = 9.14 gpm ( 0.032996 + 0.00986 ) ft/gpm 2
At this volume flow rate, the net head of the pump is 16.3 ft. Discussion
At this operating point, the flow rate is lower than that at the BEP.
14-31 Solution
We are to calculate pump efficiency and estimate the BEP conditions.
Properties
The density of water at 20oC is 998.0 kg/m3.
Analysis
(a) Pump efficiency is
ηpump =
Pump efficiency:
ρ gV H
TABLE 1 Pump performance data for water at 20oC. (1)
bhp
We show the second row of data (at V = 6.0 Lpm) as an example – the rest are calculated in a spreadsheet for convenience,
H (m) 47.5 46.2 42.5 36.2 26.2 15.0 0.0
bhp (W) 133 142 153 164 172 174 174
ηpump (%) 0.0 31.9 54.4 64.8 59.7 42.2 0.0
( 998.0 kg/m )(9.81 m/s ) ( 6.0 L/min )( 46.2 m ) ⎛ = 3
η pump
V (Lpm) 0.0 6.0 12.0 18.0 24.0 30.0 36.0
2
142 W
1 m3 ⎞ ⎛ 1 min ⎞ ⎛ N ⋅ s 2 ⎞ ⎛ W ⋅ s ⎞ ⎜ ⎟⎜ ⎟⎜ ⎟×⎜ ⎟ = 0.319 = 31.9% ⎝ 1000 L ⎠ ⎝ 60 s ⎠ ⎝ kg ⋅ m ⎠ ⎝ N ⋅ m ⎠
The results for all rows are shown in Table 1. (b) The best efficiency point (BEP) occurs at approximately the fourth row of data: V * = 18.0 Lpm, H* = 36.2 m of head, bhp* = 164. W, and ηpump* = 64.8%. Discussion
A more precise BEP could be obtained by curve-fitting the data, as in the next problem.
14-14 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-32
Solution We are to generate least-squares polynomial curve fits of a pump’s performance curves, plot the curves, and calculate the BEP. Properties
The density of water at 20oC is 998.0 kg/m3.
Analysis The efficiencies for each data point in Table P14-31 were calculated in the previous problem. We use Regression analysis to generate the least-squares fits. The equation and coefficients for H are
H = H 0 − aV 2
H 0 = 47.6643 m
a = 0.0366 m/Lpm2
The equation and coefficients for bhp are a1 = 2.37 W/Lpm
180
40
170
H (m)
160
30 150
bhp (W)
20
a = 0.0366453 m/Lpm2
Or, to 3 significant digits, H 0 = 47.7 m bhp = bhp0 + a1V + a2V 2
50
10
130
0
120 0
bhp0 = 131. W
10
a2 = -0.0317 W/Lpm2
V
The equation and coefficients for ηpump are
ηpump = ηpump,0 + a1V + a2V 2 + a3V 3 a1 = 5.87 %/Lpm
→
V * = 19.6 Lpm
At this volume flow rate, the curve-fitted expressions for H, bhp, and ηpump yield the operating conditions at the best efficiency point (to three digits each): V * = 19.6 Lpm,
H * = 33.6 m,
bhp* = 165 W ,
30
40
(Lpm) (a)
60
a3 = -0.00201 %/Lpm3
The tabulated data are plotted in Fig. 1 as symbols only. The fitted data are plotted on the same plots as lines only. The agreement is excellent. The best efficiency point is obtained by differentiating the curvefit expression for ηpump with respect to volume flow rate, and setting the derivative to zero (solving the resulting quadratic equation for V * ), dη pump = a1 + 2a2V + 3a3V 2 =0 dV
20
70
ηpump,0 = 0.152%
a2 = -0.0905 %/Lpm2
140
η * = 65.3%
50
ηpump 40 (%)
30 20 10 0 0
10
V
20
30
40
(Lpm)
(b) FIGURE 1 Pump performance curves: (a) H and bhp versus V , and (b) ηpump versus V .
Discussion This BEP is more precise than that of the previous problem because of the curve fit. The other root of the quadratic is negative − obviously not the correct choice.
14-15 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-33 Solution
For a given pump and system requirement, we are to estimate the operating point.
Assumptions
1 The flow is steady. 2 The water is at 20oC and is incompressible.
Analysis The operating point is the volume flow rate at which Hrequired = Havailable. We set the given expression for Hrequired to the curve fit expression of the previous problem, Havailable = H0 − a V 2 , and obtain V =
Operating point:
Discussion
H 0 − ( z2 − z1 ) a+b
=
47.6643 m − 10.0 m = 26.1 Lpm ( 0.0366453 + 0.0185 ) m/Lpm 2
At this operating point, the flow rate is higher than that at the BEP.
14-34 Solution We are to perform a regression analysis to estimate the shutoff head and free delivery of a pump, and then we are to determine if this pump is adequate for the system requirements. Assumptions temperature. Properties
25 20
1 The water is incompressible. 2 The water is at room o
H 15 (m) 10
3
The density of water at T = 20 C is 998.0 kg/m .
Analysis (a) We perform a regression analysis, and obtain H0 = 23.9 m and a = 0.00642 m/Lpm2. The curve fit is reasonable, as seen in Fig. 1. The shutoff head is estimated as 23.9 m of water column. At the pump’s free delivery, the net head is zero. Setting Havailable to zero in Eq. 1 gives
Free delivery: H Vmax 2 = 0 → a
Vmax =
H0 = a
5 0 0
(b) At the required operating conditions, V = 57.0 Lpm, and the net head is converted to meters of water column for analysis, H required =
( ΔP )required ρg
=
5.8 psi
( 998. kg/m )( 9.81 m/s ) 3
V
(Lpm)
FIGURE 1 Tabulated data (circles) and curve-fitted data (line) for Havailable versus V for the given pump. The filled, square data point is the required operating point.
23.9 m = 61.0 Lpm 0.00642 m/(Lpm) 2
The free delivery is estimated as 61.0 Lpm.
Required operating head:
10 20 30 40 50 60
2
⎛ 6,894.8 N/m 2 ⎞ ⎛ kg ⋅ m ⎞ ⎜ ⎟⎜ 2 ⎟ = 4.08 m psi ⎝ ⎠⎝ s ⋅ N ⎠
As seen in Fig. 1, this operating point lies above the pump performance curve. Thus, this pump is not quite adequate for the job at hand. Discussion The operating point is also very close to the pump’s free delivery, and therefore the pump efficiency would be low even if it could put out the required head.
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Chapter 14 Turbomachinery
14-35E Solution of the pump.
We are to find the units of coefficient a, write Vmax in terms of H0 and a, and calculate the operating point
Assumptions
1 The flow is steady. 2 The water is incompressible.
Analysis
(a) Solving the given expression for a gives
Coefficient a:
a=
H 0 − H available V 2
→ units of a =
ft gpm 2
(1)
(b) At the pump’s free delivery, the net head is zero. Setting Havailable to zero in the given expression gives H Vmax 2 = 0 a
Free delivery:
→
Vmax =
H0 a
(2)
(c) The operating point is obtained by matching the pump’s performance curve to the system curve. Equating these gives H available = H 0 − aV 2 = H required = ( z2 − z1 ) + bV 2
(3)
After some algebra, Eq. 3 reduces to Operating point capacity:
Voperating =
H 0 − ( z2 − z1 ) a+b
(4)
and the net pump head at the operating point is obtained by plugging Eq. 4 into the given expression, Operating point pump head:
H operating =
H 0 b + a ( z2 − z1 ) a+b
(5)
Discussion Equation 4 reveals that H0 must be greater than elevation difference (z2 – z1) in order to have a valid operating point. This agrees with our intuition, since the pump must be able to overcome the gravitational head between the tanks.
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Chapter 14 Turbomachinery
14-36 Solution
We are to calculate the operating point of a given pipe/pump system.
Assumptions temperature.
1 The water is incompressible. 2 The flow is steady since the reservoirs are large. 3 The water is at room
Properties
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Analysis
The operating point is obtained by matching the pump’s performance curve to the system curve, H available = H 0 − aV 2 = H required = ( z2 − z1 ) + bV 2
Operating point:
from which we solve for the volume flow rate (capacity) at the operating point, Voperating =
H 0 − ( z2 − z1 ) a+b
=
5.30 m − 3.52 m = 4.99 Lpm 0.0453 + 0.0261) m/Lpm 2 (
and for the net pump head at the operating point, H operating =
Discussion
H 0 b + a ( z2 − z1 ) a+b
=
( 5.30 m )( 0.0261 m ) + ( 0.0453 m )( 3.52 m ) = 4.17 m ( 0.0453 m ) + ( 0.0261 m )
The water properties ρ and μ are not needed because the system curve (Hrequired versus V ) is provided here.
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Chapter 14 Turbomachinery
14-37E Solution
For a given pump and system, we are to calculate the capacity.
Assumptions 1 The water is incompressible. 2 The flow is nearly steady since the reservoirs are large. 3 The water is at room temperature. Properties
The kinematic viscosity of water at T = 68oF is 1.055 × 10-5 ft2/s.
Analysis We apply the energy equation in head form between the inlet reservoir’s free surface (1) and the outlet reservoir’s free surface (2), H required = hpump,u =
2 2 P2 − P1 α 2 V2 − α1 V1 + + ( z2 − z1 ) + hturbine + hL ,total ρg 2g
(1)
Since both free surfaces are at atmospheric pressure, P1 = P2 = Patm, and the first term on the right side of Eq. 1 vanishes. Furthermore, since there is no flow, V1 = V2 = 0, and the second term also vanishes. There is no turbine in the control volume, so the second-to-last term is zero. Finally, the irreversible head losses are composed of both major and minor losses, but the pipe diameter is constant throughout. Equation 1 therefore reduces to 2 ⎛ L ⎞V H required = ( z2 − z1 ) + hL ,total = ( z2 − z1 ) + ⎜ f + ∑ K L ⎟ ⎝ D ⎠ 2g
(2)
The dimensionless roughness factor is ε/D = 0.0011/1.20 = 9.17 × 10-4, and the sum of all the minor loss coefficients is
∑K
L
= 0.5 + 2.0 + 6.8 + ( 3 × 0.34 ) + 1.05 = 11.37
The pump/piping system operates at conditions where the available pump head equals the required system head. Thus, we equate the given expression and Eq. 2 to find the operating point, H available = H required
→
H0 − a
π 2 D4
2 ⎛ L ⎞V V 2 = ( z2 − z1 ) + ⎜ f + ∑ KL ⎟ 16 ⎝ D ⎠ 2g
(3)
where we have written the volume flow rate in terms of average velocity through the pipe, Volume flow rate in terms of average velocity:
π D2 V = V 4
(4)
Equation 3 is an implicit equation for V since the Darcy friction factor f is a function of Reynolds number Re = ρVD/μ = VD/ν, as obtained from either the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 1.80 ft/s, from which the volume flow rate is V = 6.34 gpm. The Reynolds number is 1.67 × 104. Discussion We verify our results by comparing Havailable (given) and Hrequired (Eq. 2) at this flow rate: Havailable = 24.4 ft and Hrequired = 24.4 ft.
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Chapter 14 Turbomachinery
14-38E Solution We are to plot Hrequired and Havailable versus V , and indicate the operating point. 140
Analysis We use the equations of the previous problem, with the same constants and parameters, to generate the plot shown. The operating point is the location where the two curves intersect. The values of H and V at the operating point match those of the previous problem, as they should. Discussion A plot like this, in fact, is an alternate method of obtaining the operating point. In this case, the curve of Hrequired is fairly flat, indicating that the majority of the required pump head is attributed to elevation change, while a small fraction is attributed to major and minor head losses through the piping system.
Havailable 120 100
H 80 (ft) 60
Operating point
40 20
Hrequired
0 0
2
V
4
6
8
(gpm)
14-39E Solution We are to re-calculate volume flow rate for a piping system with a much longer pipe, and we are to compare with the previous results. Analysis All assumptions, properties, dimensions, and parameters are identical to those of the previous problem, except that total pipe length L is longer. We repeat the calculations and find that V = 1.68 ft/s, from which the volume flow rate is V = 5.93 gpm, and the net head of the pump is 37.0 ft. The Reynolds number for the flow in the pipe is 1.56 × 104. The volume flow rate has decreased by about 6.5%. Discussion The decrease in volume flow rate is smaller than we may have suspected. This is because the majority of the pump work goes into raising the elevation of the water. In addition, as seen in the plot from the previous problem, the pump performance curve is quite steep near these flow rates – a significant change in required net head leads to a much less significant change in volume flow rate.
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Chapter 14 Turbomachinery
14-40E Solution We are to perform a regression analysis to translate tabulated pump performance data into an analytical expression, and then use this expression to predict the volume flow rate through a piping system. Assumptions 1 The water is incompressible. 2 The flow is nearly steady since the reservoirs are large. 3 The water is at room temperature. Properties For water at T = 68oF, μ = 6.572 × 10-4 lbm/ft⋅s, and ρ = 3 62.31 lbm/ft , from which ν = 1.055 × 10-5 ft2/s.
45
Data points
40 35 30 H 25 (ft) 20
Curve fit
15 10
40
5 0
30
Havailable
0
Hrequired
H (ft) 20
10
V
15
20
25
(gpm)
FIGURE 1 Tabulated data (symbols) and curve-fitted data (line) for Havailable versus V for the proposed pump.
Operating point
10
5
0 0
5
10
V
15
20
25
(gpm)
FIGURE 2 Havailable and Hrequired versus V for a piping system with pump; the operating point is also indicated, where the two curves meet. Analysis (a) We perform a regression analysis, and obtain H0 = 38.15 ft and a = 0.06599 ft/gpm2. The curve fit is very good, as seen in Fig. 1.
(b) We repeat the calculations of Problem 14-37 with the new pump performance coefficients, and find that V = 3.29 ft/s, from which the volume flow rate is V = 11.6 gpm, and the net head of the pump is 29.3 ft. The Reynolds number for the flow in the pipe is 3.05 × 104. The volume flow rate has increased by about 83%. Paul is correct – this pump performs much better, nearly doubling the flow rate. (c) A plot of net head versus volume flow rate is shown in Fig. 2. Discussion
This pump is more appropriate for the piping system at hand.
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Chapter 14 Turbomachinery
14-41 Solution
For a given pump and system, we are to calculate the capacity.
Assumptions 1 The water is incompressible. 2 The flow is nearly steady since the reservoirs are large. 3 The water is at room temperature. Properties
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Analysis We apply the energy equation in head form between the inlet reservoir’s free surface (1) and the outlet reservoir’s free surface (2), H required = hpump,u =
2 2 P2 − P1 α 2 V2 − α1 V1 + + ( z2 − z1 ) + hturbine + hL ,total ρg 2g
(1)
Since both free surfaces are at atmospheric pressure, P1 = P2 = Patm, and the first term on the right side of Eq. 1 vanishes. Furthermore, since there is no flow, V1 = V2 = 0, and the second term also vanishes. There is no turbine in the control volume, so the second-to-last term is zero. Finally, the irreversible head losses are composed of both major and minor losses, but the pipe diameter is constant throughout. Equation 1 therefore reduces to 2 ⎛ L ⎞V H required = ( z2 − z1 ) + hL ,total = ( z2 − z1 ) + ⎜ f + ∑ KL ⎟ ⎝ D ⎠ 2g
(2)
The dimensionless roughness factor is ε 0.25 mm ⎛ 1 cm ⎞ = = 0.0123 D 2.03 cm ⎜⎝ 10 mm ⎟⎠
The sum of all the minor loss coefficients is
∑K
L
= 0.5 + 17.5 + ( 5 × 0.92 ) + 1.05 = 23.65
The pump/piping system operates at conditions where the available pump head equals the required system head. Thus, we equate the given expression and Eq. 2 to find the operating point, H available = H required
→
H0 − a
π 2 D4
2 ⎛ L ⎞V V 2 = ( z2 − z1 ) + ⎜ f + ∑ KL ⎟ 16 ⎝ D ⎠ 2g
(3)
where we have written the volume flow rate in terms of average velocity through the pipe,
π D2 V = V 4 Equation 3 is an implicit equation for V since the Darcy friction factor f is a function of Reynolds number Re = ρVD/μ, as obtained from either the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 0.59603 ≈ 0.596 m/s, from which the volume flow rate is V = 11.6 Lpm. The Reynolds number is 1.21 × 104. Discussion We verify our results by comparing Havailable (given) and Hrequired (Eq. 2) at this flow rate: Havailable = 15.3 m and Hrequired = 15.3 m.
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Chapter 14 Turbomachinery
14-42 Solution We are to plot Hrequired and Havailable versus V , and indicate the operating point. Analysis We use the equations of the previous problem, with the same constants and parameters, to generate the plot shown. The operating point is the location where the two curves intersect. The values of H and V at the operating point match those of the previous problem, as they should. Discussion A plot like this, in fact, is an alternate method of obtaining the operating point.
30
Havailable
25 20
Operating point
H (m) 15 10
Hrequired
5 0 0
5
10
V (Lpm)
15
20
14-43 Solution We are to re-calculate volume flow rate for a piping system with a smaller elevation difference, and we are to compare with the previous results. Analysis All assumptions, properties, dimensions, and parameters are identical to those of the previous problem, except that the elevation difference between reservoir surfaces (z2 – z1) is smaller. We repeat the calculations and find that V = 0.682 m/s, from which the volume flow rate is V = 13.2 Lpm, and the net head of the pump is 12.5 m. The Reynolds number for the flow in the pipe is 1.38 × 104. The volume flow rate has increased by about 14%. Discussion The increase in volume flow rate is modest. This is because only about half of the pump work goes into raising the elevation of the water – the other half goes into overcoming irreversible losses.
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Chapter 14 Turbomachinery
14-44 Solution We are to perform a regression analysis to translate tabulated pump performance data into an analytical expression, and then use this expression to predict the volume flow rate through a piping system. Assumptions 1 The water is incompressible. 2 The flow is nearly steady since the reservoirs are large. 3 The water is at room temperature. o
Properties The density and viscosity of water at T = 20 C are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
50
Data points 40
H (m)
30
Curve fit 20 10
50
0 0
40
H 30 (m) 20
10
20
30
Havailable
V
Operating point
FIGURE 1 Tabulated data (symbols) and curve-fitted data (line) for Havailable versus V for the proposed pump.
10
(Lpm)
Hrequired
0 0
5
10
V
15
20
25
30
(Lpm)
FIGURE 2 Havailable and Hrequired versus V for a piping system with pump; the operating point is also indicated, where the two curves meet. Analysis (a) We perform a regression analysis, and obtain H0 = 47.6 m and a = 0.05119 m/Lpm2. The curve fit is reasonable, as seen in Fig. 1.
(b) We repeat the calculations of Problem 14-41 with the new pump performance coefficients, and find that V = 1.00 m/s, from which the volume flow rate is V = 19.5 Lpm, and the net head of the pump is 28.3 m. The Reynolds number for the flow in the pipe is 2.03 × 104. The volume flow rate has increased by about 69%. April’s goal has not been reached. She will need to search for an even stronger pump. (c) A plot of net head versus volume flow rate is shown in Fig. 2. Discussion As is apparent from Fig. 2, the required net head increases rapidly with increasing volume flow rate. Thus, doubling the flow rate would require a significantly heftier pump.
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Chapter 14 Turbomachinery
14-45 Solution
We are to calculate the volume flow rate when the pipe diameter of a piping/pump system is doubled.
Analysis The analysis is identical to that of Problem 14-41 except for the diameter change. The calculations yield V = 0.19869 ≈ 0.199 m/s, from which the volume flow rate is V = 15.4 Lpm, and the net head of the pump is 8.25 m. The Reynolds number for the flow in the pipe is 8.03 × 103. The volume flow rate has increased by about 33%. This agrees with our intuition since irreversible head losses go down significantly by increasing pipe diameter. Discussion The gain in volume flow rate is significant because the irreversible head losses contribute to about half of the total pump head requirement in the original problem.
14-46 Solution We are to compare Reynolds numbers for a pipe flow system – the second case having a pipe diameter twice that of the first case. Properties
The density and viscosity of water at T = 20oC are 998.0 kg/m3 and 1.002 × 10-3 kg/m⋅s respectively.
Analysis
From the results of the two problems, the Reynolds number of the first case is
Case 1 (D = 2.03 cm): Re =
3 ρVD ( 998 kg/m ) ( 0.59603 m/s )( 0.0203 m ) = = 1.21× 104 μ 1.002 × 10-3 kg/m3
and that of the second case is
Case 2 (D = 4.06 cm):
3 ρVD ( 998 kg/m ) ( 0.19869 m/s )( 0.0406 m ) = = 0.803 × 104 Re = μ 1.002 × 10-3 kg/m3
Thus, the Reynolds number of the larger diameter pipe is smaller than that of the smaller diameter pipe. This may be somewhat surprising, but since average pipe velocity scales as the inverse of pipe diameter squared, Reynolds number increases linearly with pipe diameter due to the D in the numerator, but decreases quadratically with pipe diameter due to the V in the numerator. The net effect is a decrease in Re with pipe diameter when V is the same. In this problem, V increases somewhat as the diameter is doubled, but not enough to increase the Reynolds number. Discussion At first glance, most people would think that Reynolds number increases as both diameter and volume flow rate increase, but this is not always the case.
14-47 Solution
We are to compare the volume flow rate in a piping system with and without accounting for minor losses.
Analysis The analysis is identical to that of Problem 14-41, except we ignore all the minor losses. The calculations yield V = 0.604 m/s, from which the volume flow rate is V = 11.7 Lpm, and the net head of the pump is 15.1 m. The Reynolds number for the flow in the pipe is 1.22 × 104. The volume flow rate has increased by about 1.3%. Thus, minor losses are nearly negligible in this calculation. This agrees with our intuition since the pipe is very long. Discussion Since the Colebrook equation is accurate to at most 5%, a 1.3% change is well within the error. Nevertheless, it is not excessively difficult to include the minor losses, especially when solving the problem on a computer.
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Chapter 14 Turbomachinery
14-48 Solution We are to examine how increasing (z2 – z1) affects the volume flow rate of water pumped by the water pump. Assumptions 1 The flow at any instant of time is still considered quasisteady, since the surface level of the upper reservoir rises very slowly. 2 The minor losses, dimensions, etc., fluid properties, and all other assumptions are identical to those of Problem 14-41 except for the elevation difference (z2 – z1).
14 12
V (Lpm)
10 8 6
4 Analysis We repeat the calculations of Problem 14-41 for several values of (z2 – z1), ranging from 0 to H0, the shutoff head of the pump, 2 since above the shutoff head, the pump cannot overcome the elevation 0 difference. The volume flow rate is zero at the shutoff head of the pump. The data are plotted here. As expected, the volume flow rate 0 5 10 15 20 25 decreases as (z2 – z1) increases, starting at a maximum flow rate of about z2 – z1 (m) 14.1 Lpm when there is no elevation difference, and reaching zero (no flow) when (z2 – z1) = H0 = 24.4 m. The curve is not linear, since neither the Darcy friction factor nor the pump performance curve are linear. If (z2 – z1) were increased beyond H0, the pump would not be able to handle the elevation difference. Despite its valiant efforts, with blades spinning as hard as they could, the water would flow backwards through the pump.
Discussion You may wish to think of the backward-flow through the pump as a case in which the pump efficiency is negative. In fact, at (z2 – z1) = H0, the pump could be replaced by a closed valve to keep the water from draining from the upper reservoir to the lower reservoir.
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Chapter 14 Turbomachinery
14-49E Solution
We are to estimate the operating point of a given fan and duct system.
Assumptions 1 The flow is steady. 2 The concentration of contaminants in the air is low; the fluid properties are those of air alone. 3 The air is at standard temperature and pressure (STP), and is incompressible. 4 The air flowing in the duct is turbulent with α = 1.05. Properties For air at STP (T = 77oF, P = 14.696 psi = 2116.2 lbf/ft2), μ = 1.242 × 10-5 lbm/ft⋅s, ρ = 0.07392 lbm/ft3, and ν = 1.681 × 10-4 ft2/s. The density of water at STP (for conversion to inches of water head) is 62.24 lbm/ft3. Analysis We apply the steady energy equation along a streamline from point 1 in the stagnant air region in the room to point 2 at the duct outlet, H required =
Required net head:
2 2 P2 − P1 α 2V2 − α1 V1 + ( z2 − z1 ) + hL ,total + ρg 2g
(1)
At point 1, P1 is equal to Patm, and at point 2, P2 is also equal to Patm since the jet discharges into the outside air on the roof of the building. Thus the pressure terms cancel out in Eq. 1. We ignore the air speed at point 1 since it is chosen (wisely) far enough away from the hood inlet so that the air is nearly stagnant. Finally, the elevation difference is neglected for gases. Equation 1 reduces to H required =
α 2V2 2 2g
+ hL ,total
(2)
The total head loss in Eq. 2 is a combination of major and minor losses, and depends on volume flow rate. Since the duct diameter is constant, 2 ⎛ L ⎞V hL ,total = ⎜ f + ∑ KL ⎟ ⎝ D ⎠ 2g
(3)
2 L ⎛ ⎞V H required = ⎜ α 2 + f + ∑ K L ⎟ D ⎝ ⎠ 2g
(4)
Total irreversible head loss: The required net head of the fan is thus
To find the operating point, we equate Havailable and Hrequired, being careful to keep consistent units. Note that the required head in Eq. 4 is expressed naturally in units of equivalent column height of the pumped fluid, which is air in this case. However, the available net head (given) is in terms of equivalent water column height. We convert constants H0 and a to inches of air column for consistency by multiplying by the ratio of water density to air density, H 0, inch water ρ water = H 0, inch air ρ air
→
H 0, inch air = H 0, inch water
ρ water ρ air
and similarly, a( inch air ) / SCFM2 = a(inch water ) / SCFM2
ρ water ρair
We re-write the given expression in terms of average duct velocity rather than volume flow rate, H available = H 0 − a
Available net head:
π 2 D4 16
V2
(5)
again taking care to keep consistent units. Equating Eqs. 4 and 5 yields
Operating point:
H available = H required
→
H0 − a
π 2 D4
2 L ⎛ ⎞V V 2 = ⎜α 2 + f + ∑ KL ⎟ 16 D ⎝ ⎠ 2g
(6)
The dimensionless roughness factor is ε/D = 0.0059/9.06 = 6.52 × 10-4, and the sum of all the minor loss coefficients is
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Chapter 14 Turbomachinery
Minor losses:
∑K
L
= 4.6 + ( 3 × 0.21) + 1.8 = 7.03
Note that there is no minor loss associated with the exhaust, since point 2 is taken at the exit plane of the duct, and does not include irreversible losses associated with the turbulent jet. Equation 6 is an implicit equation for V since the Darcy friction factor is a function of Reynolds number Re = ρVD/μ = VD/ν, as obtained from either the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 16.8 ft/s, from which the volume flow rate is V = 452 SCFM. The Reynolds number is 7.63 × 104. Discussion We verify our results by comparing Havailable (Eq. 1) and Hrequired (Eq. 5) at this flow rate: Havailable = 0.566 inches of water and Hrequired = 0.566 inches of water.
14-50E Solution
We are to calculate the value of KL, damper such that the volume flow rate through the duct decreases by 50%.
Analysis All assumptions and properties are the same as those of the previous problem. We set the volume flow rate to V = 226 SCFM, one-half of the previous result, and solve for KL, damper. The result is KL, damper = 112, significantly higher than the value of 1.8 for the fully open case. Discussion Because of the nonlinearity of the problem, we cannot simply double the damper’s loss coefficient in order to decrease the flow rate by a factor of two. Indeed, the minor loss coefficient must be increased by a factor of more than 60. If a computer was used for the calculations of the previous problem, the solution here is most easily obtained by trial and error.
14-51E Solution We are to estimate the volume flow rate at the operating point without accounting for minor losses, and then we are to compare with the previous results. Analysis All assumptions and properties are the same as those of Problem 14-49, except that we ignore all minor losses (we set ΣKL = 0). The resulting volume flow rate at the operating point is V = 503 SCFM, approximately 11% higher than for the case with minor losses taken into account. In this problem, minor losses are indeed “minor”, although they are not negligible. We should not be surprised at this result, since there are several minor losses, and the duct is not extremely long (L/D is only 45.0). Discussion An error of 11% may be acceptable in this type of problem. However, since it is not difficult to account for minor losses, especially if the calculations are performed on a computer, it is wise not to ignore these terms.
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Chapter 14 Turbomachinery
14-52 Solution
We are to estimate the operating point of a given fan and duct system.
Assumptions 1 The flow is steady and incompressible. 2 The concentration of contaminants is low; the fluid properties are those of air alone. 3 The air is at 25oC and 101,300 Pa. 4 The air flowing in the duct is turbulent with α = 1.05. Properties For air at 25oC, μ = 1.849 × 10-5 kg/m⋅s, ρ = 1.184 kg/m3, and ν = 1.562 × 10-5 m2/s. The density of water at STP (for conversion to water head) is 997.0 kg/m3. Analysis We apply the steady energy equation along a streamline from point 1 in the stagnant air region in the room to point 2 at the duct outlet, H required =
2 2 P2 − P1 α 2V2 − α1 V1 + ( z2 − z1 ) + hL ,total + ρg 2g
(1)
P1 is equal to Patm, and P2 is also equal to Patm since the jet discharges into outside air on the roof of the building. Thus the pressure terms cancel out in Eq. 1. We ignore the air speed at point 1 since it is chosen (wisely) far enough away from the hood inlet so that the air is nearly stagnant. Finally, the elevation difference is neglected for gases. Equation 1 reduces to H required =
Required net head:
α 2V2 2 2g
+ hL ,total
(2)
The total head loss in Eq. 2 is a combination of major and minor losses. Since the duct diameter is constant, 2 ⎛ L ⎞V hL ,total = ⎜ f + ∑ K L ⎟ ⎝ D ⎠ 2g
(3)
2 L ⎛ ⎞V H required = ⎜ α 2 + f + ∑ KL ⎟ D ⎝ ⎠ 2g
(4)
Total irreversible head loss: The required net head of the fan is thus
To find the operating point, we equate Havailable and Hrequired, being careful to keep consistent units. Note that the required head in Eq. 4 is expressed naturally in units of equivalent column height of the pumped fluid, which is air in this case. However, the available net head (given) is in terms of equivalent water column height. We convert constants H0 and a in Eq. 1 to mm of air column for consistency by multiplying by the ratio of water density to air density, H 0, mm water ρ water = H 0, mm air ρ air
→
H 0, mm air = H 0, mm water
ρ water ρair
and
a( mm air ) / LPM 2 = a( mm water ) / LPM2
ρ water ρ air
We re-write the given expression in terms of average duct velocity rather than volume flow rate, H available = H 0 − a
π 2 D4
V2
(5)
2 L ⎛ ⎞V V 2 = ⎜α 2 + f + ∑ KL ⎟ 16 D ⎝ ⎠ 2g
(6)
Available net head:
16
Equating Eqs. 4 and 5 yields the operating point, H available = H required
→
H0 − a
π 2 D4
The dimensionless roughness factor is ε/D = 0.15/150 = 1.00 × 10-3, and the sum of all the minor loss coefficients is ∑ K L = 3.3 + ( 3 × 0.21) + 1.8 + 0.36 + 6.6 = 12.69 . Note that there is no minor loss associated with the exhaust, since point 2 is at the exit plane of the duct, and does not include irreversible losses associated with the turbulent jet. Equation 6 is an implicit equation for V since the Darcy friction factor is a function of Reynolds number Re = ρVD/μ = VD/ν, as obtained from the Moody chart or the Colebrook equation. The solution can be obtained by an iterative method, or through use of a mathematical equation solver like EES. The result is V = 6.71 m/s, from which the volume flow rate is V = 7090 Lpm. Discussion We verify our results by comparing Havailable (given) and Hrequired (Eq. 5) at this flow rate: Havailable = 47.4 mm of water and Hrequired = 47.4 mm of water, both of which are equivalent to 40.0 m of air column.
14-29 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-53 Solution We are to plot Hrequired and Havailable versus V , and indicate the operating point. Analysis We use the equations of the previous problem, with the same constants and parameters, to generate the plot shown. The operating point is the location where the two curves intersect. The values of H and V at the operating point match those of the previous problem, as they should.
100 80
Havailable Operating point
H 60 (mm) 40
Discussion A plot like this, in fact, is an alternate method of obtaining the operating point. The operating point is at a volume flow rate near the center of the plot, indicating that the fan efficiency is probably reasonably high.
20
Hrequired 0 0
5000
10000
V (Lpm)
15000
14-54 Solution We are to estimate the volume flow rate at the operating point without accounting for minor losses, and then we are to compare with the previous results. Analysis All assumptions and properties are the same as those of Problem 14-52, except that we ignore all minor losses (we set ΣKL = 0). The resulting volume flow rate at the operating point is V = 10,900 Lpm (to three significant digits), approximately 54% higher than for the case with minor losses taken into account. In this problem, minor losses are not “minor”, and are by no means negligible. Even though the duct is fairly long (L/D is about 163), the minor losses are large, especially those through the damper and the one-way valve. Discussion An error of 54% is not acceptable in this type of problem. Furthermore, since it is not difficult to account for minor losses, especially if the calculations are performed on a computer, it is wise not to ignore these terms.
14-55 Solution
We are to calculate pressure at two locations in a blocked duct system.
Assumptions 1 The flow is steady. 2 The concentration of contaminants in the air is low; the fluid properties are those of air alone. 3 The air is at standard temperature and pressure (STP: 25oC and 101,300 Pa), and is incompressible. Properties
The density of water at 25oC is 997.0 kg/m3.
Analysis Since the air is completely blocked by the one-way valve, there is no flow. Thus, there are no major or minor losses – just a pressure gain across the fan. Furthermore, the fan is operating at its shutoff head conditions. Since the pressure in the room is atmospheric, the gage pressure anywhere in the stagnant air region in the duct between the fan and the one-way valve is therefore equal to H0 = 60.0 mm of water column. We convert to pascals as follows:
Gage pressure at both locations:
⎛ N ⋅ s 2 ⎞ ⎛ Pa ⋅ m 2 ⎞ kg ⎞ ⎛ m⎞ ⎛ Pgage = ρ water gH 0 = ⎜ 998.0 3 ⎟ ⎜ 9.81 2 ⎟ ( 0.060 m ) ⎜ ⎟⎜ ⎟ = 587 Pa m ⎠⎝ s ⎠ ⎝ ⎝ kg ⋅ m ⎠ ⎝ N ⎠
Thus, at either location, the gage pressure is 60.0 mm of water column, or 587 Pa. Discussion The answer depends only on the shutoff head of the fan – duct diameter, minor losses, etc, are irrelevant for this case since there is no flow. The fan should not be run for long time periods under these conditions, or it may burn out.
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Chapter 14 Turbomachinery 14-56E Solution For a given pump and piping system we are to estimate the maximum volume flow rate that can be pumped without cavitation. Assumptions
1 The flow is steady. 2 The water is incompressible. 3 The flow is turbulent.
Properties Patm = 14.696 psi = 2116.2 lbf/ft2. For water at T = 77oF, μ = 6.002 × 10-4 lbm/(ft⋅s), ρ = 62.24 lbm/ft3, and Pv = 66.19 lbf/ft2. Analysis We apply the steady energy equation in head form along a streamline from point 1 at the reservoir surface to point 2 at the pump inlet,
P1 αV2 P αV2 + 1 1 + z1 + hpump,u = 2 + 2 2 + z2 + hturbine + hL ,total 2g 2g ρg ρg
(1)
60
In Eq. 1 we have ignored the water speed at the reservoir surface (V1 ≈ 0). There is no turbine in the piping system. Also, although there is a pump in the system, there is no pump between point 1 and point 2; hence the pump head term also drops out. We solve Eq. 1 for P2/(ρg), which is the pump inlet pressure expressed as a head,
50
NPSH (ft)
40 30
P αV2 P2 = atm + ( z1 − z2 ) − 2 2 − hL ,total 2g ρg ρg
Available NPSH
No cavitation
(2) 20
Note that in Eq. 2, we have recognized that P1 = Patm since the reservoir surface is exposed to atmospheric pressure. The available net positive suction head at the pump inlet is obtained from Eq. 14-8. After substitution of Eq. 2, approximating α2 ≈ 1, we get an expression for the available NPSH:
10
Required NPSH
0 0 10 20 30 40 50 60 70
P − Pv NPSH = atm + ( z1 − z2 ) − hL ,total ρg
(3)
V (gpm)
Since we know Patm, Pv, and the elevation difference, all that remains is to estimate the total irreversible head loss through the piping system from the reservoir surface (1) to the pipe inlet (2), which depends on volume flow rate. Since the pipe diameter is constant, the total irreversible head loss becomes 2 ⎛ L ⎞V hL ,total = ⎜ f + ∑ KL ⎟ ⎝ D ⎠ 2g
(4)
The rest of the problem is most easily solved on a computer. For a given volume flow rate, we calculate speed V and Reynolds number Re. From Re and the known pipe roughness, we use the Moody chart (or the Colebrook equation) to obtain friction factor f. The sum of all the minor loss coefficients is
∑K
L
= 0.5 + 0.3 + 6.0 = 6.8
(5)
where we have not included the minor losses downstream of the pump, since they are irrelevant to the present analysis. We make one calculation by hand for illustrative purposes. At V = 40.0 gpm, the average speed of water through the pipe is V=
4 ( 40.0 gal/min ) ⎛ 231 in 3 4V V = = ⎜ 2 A π D2 π (1.2 in ) ⎝ gal
⎞ ⎛ 1 min ⎞⎛ 1 ft ⎞ ⎟⎜ ⎟⎜ ⎟ = 11.35 ft/s ⎠ ⎝ 60 s ⎠⎝ 12 in ⎠
(6)
which produces a Reynolds number of Re = ρVD/μ = 1.17 × 105. At this Reynolds number, and with roughness factor ε/D = 0, the Colebrook equation yields f = 0.0174. After substitution of the given variables along with f, D, L, and Eqs. 4, 5, and 6 into Eq. 3, we get
14-31 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery NPSH =
( 2116.2 − 66.19 ) lbf/ft 2
2
( 62.24 lbm/ft )( 32.174 ft/s ) 3
12.0 ft ⎛ 32.174 lbm ft ⎞ ⎛ ⎞ (11.35 ft/s ) + 20.0 ft − ⎜ 0.0174 + 6.8 ⎟ = 35.1 ft ⎜ ⎟ 2 0.10 ft s lbf ⎝ ⎠ ⎝ ⎠ 2 32.174 ft/s 2
2
(
)
(7)
The required net positive suction head is obtained from the given expression. At our example flow rate of 40.0 gpm we see that NPSHrequired is about 9.6 ft. Since the actual NPSH is much higher than this, we need not worry about cavitation at this flow rate. We use a spreadsheet to calculate NPSH as a function of volume flow rate, and the results are plotted. It is clear from the plot that cavitation occurs at flow rates above about 56 gallons per minute. Discussion
NPSHrequired rises with volume flow rate, but the actual or available NPSH decreases with volume flow rate.
14-57E Solution
We are to calculate the volume flow rate below which cavitation in a pump is avoided.
Assumptions
1 The flow is steady. 2 The water is incompressible.
Properties Patm = 14.696 psi = 2116.2 lbf/ft2. For water at T = 150oF, μ = 2.889 × 10-4 lbm/ft⋅s, ρ = 61.19 lbm/ft3, and Pv = 536.0 lbf/ft2. Analysis The procedure is identical to that of the previous problem, except for the water properties. The calculations predict that the pump cavitates at volume flow rates greater than about 53 gpm. This is somewhat lower than the result of the previous problem, as expected, since cavitation occurs more readily in warmer water. Discussion Note that NPSHrequired does not depend on water temperature, but the actual or available NPSH decreases with temperature.
14-32 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-58 Solution For a given pump and piping system we are to estimate the maximum volume flow rate that can be pumped without cavitation. Assumptions
1 The flow is steady. 2 The water is incompressible. 3 The flow is turbulent.
Properties Standard atmospheric pressure is Patm = 101.3 kPa. For water at T = 25oC, ρ = 997.0 kg/m3, μ = 8.91×10-4 kg/m⋅s, and Pv = 3.169 kPa. Analysis We apply the steady energy equation in head form along a streamline from point 1 at the reservoir surface to point 2 at the pump inlet, P1 αV2 P αV2 + 1 1 + z1 + hpump,u = 2 + 2 2 + z2 + hturbine + hL ,total 2g 2g ρg ρg
Energy equation:
(1)
In Eq. 1 we have ignored the water speed at the reservoir surface (V1 ≈ 0). There is no turbine in the piping system. Also, although there is a pump in the system, there is no pump between point 1 and point 2; hence the pump head term also drops out. We solve Eq. 1 for P2/(ρg), which is the pump inlet pressure expressed as a head,
P αV2 P2 = atm + ( z1 − z2 ) − 2 2 − hL ,total 2g ρg ρg
Pump inlet pressure head:
(2)
Note that in Eq. 2, we have recognized that P1 = Patm since the reservoir surface is exposed to atmospheric pressure. The available net positive suction head at the pump inlet is obtained from Eq. 14-8. After substitution of Eq. 2, and approximating α2 as 1.0, we get 9 Available Patm − Pv + ( z1 − z2 ) − hL ,total Available NPSH: NPSH = (3) NPSH ρg 8 Since we know Patm, Pv, and the elevation difference, all that remains is to estimate the total irreversible head loss through the piping system from the reservoir surface (1) to the pipe inlet (2), which depends on volume flow rate. Since the pipe diameter is constant, the total irreversible head loss is 2 ⎛ L ⎞V hL ,total = ⎜ f + ∑ K L ⎟ ⎝ D ⎠ 2g
L
= 0.85 + 0.3 = 1.15
NPSH (m) 6 No cavitation
5
(4)
The rest of the problem is most easily solved on a computer. For a given volume flow rate, we calculate speed V and Reynolds number Re. From Re and the known pipe roughness, we use the Moody chart (or the Colebrook equation) to obtain friction factor f. The sum of all the minor loss coefficients is
∑K
7
Required NPSH
4 3 2 0
(5)
10 20 30 40 50 60 70
V (Lpm)
where we have not included the minor losses downstream of the pump, since they are irrelevant to the present analysis. We make one calculation by hand for illustrative purposes. At V = 40.0 Lpm, the average speed of water through the pipe is V=
4 ( 40.0 L/min ) ⎛ 1 m3 ⎞ ⎛ 1 min ⎞ 4V V = = ⎟⎜ ⎟ = 1.474 m/s 2 ⎜ A π D2 π ( 0.024 m ) ⎝ 1000 L ⎠ ⎝ 60 s ⎠
(6)
which produces a Reynolds number of Re = ρVD/μ = 3.96 × 104. At this Reynolds number, and with roughness factor ε/D = 0, the Colebrook equation yields f = 0.0220. After substitution of the given variables, along with f, D, L, and Eqs. 4, 5, and 6 into Eq. 3, we calculate the available NPSH,
14-33 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery NPSH =
(101,300 − 3,169 ) N/m 2
( 997.0 kg/m )( 9.81 m/s ) 3
2
2.8 m ⎛ kg m ⎞ ⎛ ⎞ (1.474 m/s ) = 7.42 m ⎜ s 2 N ⎟ − 2.2 m − ⎜ 0.0220 0.024 m + 1.15 ⎟ ⎝ ⎠ ⎝ ⎠ 2 9.81 m/s 2 2
(
)
(7)
The required net positive suction head is obtained from the given expression. At our example flow rate of 40.0 Lpm we see that NPSHrequired is about 4.28 m. Since the actual NPSH is higher than this, the pump does not cavitate at this flow rate. We use a spreadsheet to calculate NPSH as a function of volume flow rate, and the results are plotted. It is clear from this plot that cavitation occurs at flow rates above 60.5 liters per minute. Discussion
NPSHrequired rises with volume flow rate, but the actual or available NPSH decreases with volume flow rate.
14-59 Solution We are to calculate the volume flow rate below which cavitation in a pump is avoided, at two temperatures. Assumptions
1 The flow is steady. 2 The water is incompressible.
Properties Standard atmospheric pressure is Patm = 101.3 kPa. For water at T = 80oC, ρ = 971.9 kg/m3, μ = 3.55×10-4 kg/m⋅s, and Pv = 47.35 kPa. At T = 90oC, ρ = 965.3 kg/m3, μ = 3.15×10-4 kg/m⋅s, and Pv = 70.11 kPa. Analysis The procedure is identical to that of the previous problem, except for the water properties. The calculations predict that at T = 80oC, the pump cavitates at volume flow rates greater than 28.0 Lpm. This is substantially lower than the result of the previous problem, as expected, since cavitation occurs more readily in warmer water. At 90oC, the vapor pressure is very high since the water is near boiling (at atmospheric pressure, water boils at 100oC). For this case, the curves of NPSHavailable and NPSHrequired do not cross at all as seen in the plot, implying that the pump cavitates at any flow rate when T = 90oC. Discussion Note that NPSHrequired does not depend on water temperature, but the actual or available NPSH decreases with temperature.
9 8 7 6 NPSH 5 (m) 4
Required NPSH
3
Available NPSH
2 1 0 0
10 20 30 40 50 60 70 V (Lpm)
14-60 Solution We are to calculate the volume flow rate below which cavitation in a pump is avoided, and compare with a previous case with a smaller pipe diameter. Assumptions
1 The flow is steady. 2 The water is incompressible.
Properties Standard atmospheric pressure is Patm = 101.3 kPa. For water at T = 25oC, ρ = 997.0 kg/m3, μ = 8.91×10-4 kg/m⋅s, and Pv = 3.169 kPa. Analysis The analysis is identical to that of Problem 14-58, except that the pipe diameter is 48.0 mm instead of 24.0 mm. Compared to problem 14-58, at a given volume flow rate, the average speed through the pipe decreases by a factor of four since the pipe area increases by a factor of four. The Reynolds number goes down by a factor of two, but the flow is still turbulent (At our sample flow rate of 40.0 Lpm, Re = 1.98 × 104). For a smooth pipe at this Reynolds number, f = 0.0260. and the available NPSH is 7.81 m, slightly higher than the previous case with the smaller diameter pipe. After repeating the calculations at several flow rates, we find that the pump cavitates at V = 65.5 Lpm. This represents an increase of about 8.3%. Cavitation occurs at a higher volume flow rate when the pipe diameter is increased because the irreversible head losses in the piping system upstream of the pump are decreased. Discussion If a computer program like EES was used for Problem 14-58, it is a trivial matter to change the pipe diameter and re-do the calculations.
14-34 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-61 Solution We are to calculate the combined shutoff head and free delivery for two pumps in series, and discuss why the weaker pump should be shut off and bypassed above some flow rate. Assumptions
1 The water is incompressible. 2 The flow is steady.
Analysis The pump performance curves for both pumps individually and for their combination in series are plotted in Fig. 1. At zero flow rate, the shutoff head of the two pumps in series is equal to the sum of their individual shutoff heads: H0,combined = H0,1 + H0,2 = 5.30 m + 7.80 m. Thus the combined shutoff head is 13.1 m. At free delivery conditions (zero net head), the free delivery of the two pumps in series is limited to that of the stronger pump, in this case Pump 2: Vmax,combined = , V ) = 15.0 Lpm. Thus the combined free delivery is MAX( V max,1
14
Combined net head 12 10
Pump 2
6 4
Pump 1
max,2
2
15.0 Lpm. As volume flow rate increases, the combined net pump head is equal to the sum of the net pump heads of the individual pumps, as seen on the plot. However, the free delivery of Pump 1 is 0 = H 0 − aV 2
Bypass Pump 1
H 8 (m)
→
Vmax =
H0 = a
0 0
5 10 V ( Lpm )
5.30 m = 11.0 Lpm 0.0438 m/(Lpm) 2
15
Above this flow rate, Pump 1 no longer contributes to the flow. In fact, it becomes a liability to the system since its net head would be negative at flow rates above 11.0 Lpm. For this reason, it is wise to shut off and bypass Pump 1 above 11.0 Lpm. Discussion The free delivery of Pump 2 is 15.0 Lpm. The free delivery of the two-pump system is no better than that of Pump 2 alone, since Pump 1 is shut off and bypassed at flow rates above 11.0 Lpm.
14-62 Solution We are to calculate the free delivery and shutoff head for two pumps in parallel, and discuss why the weaker pump should be shut off and bypassed above some net head. 8 Assumptions 1 The water is incompressible. 2 The flow is steady. Analysis The pump performance curves for both pumps and for their combination in parallel are plotted. At zero flow rate, the shutoff head of the two pumps in parallel is equal to that of the stronger pump, in this case Pump 2: H0,combined = MAX(H0,1, H0,2) = 7.80 m. Thus the combined shutoff head is 7.80 m. At free delivery conditions (zero net head), the free delivery of the two pumps in parallel is the sum of their individual free deliveries: Vmax,combined = Vmax,1 + Vmax,2 = 11.0 Lpm + 15.0
Lpm. Thus the combined free delivery is 26.0 Lpm. As net head increases, the combined capacity is equal to the sum of the capacities of the individual pumps, as seen on Fig. 1. However, the shutoff head of Pump 1 (5.30 m) is lower than that of Pump 2 (7.80 m). Above the shutoff head of Pump 1, that pump no longer contributes to the flow. In fact, it becomes a liability to the system since it cannot sustain such a high head. If not shut off and blocked, the volume flow rate through Pump 1 would be negative at net heads above 7.80 m. For this reason, it is wise to shut off and block Pump 1 above 7.80 m.
7
Block Pump 1
6
H 5 (m) 4
Combined capacity
3 2
Pump 2 1
Pump 1 0 0
5
10 15 20 V ( Lpm )
25
Discussion The shutoff head of Pump 2 is 7.80 m. The shutoff head of the two-pump system is no better than that of Pump 2 alone, since Pump 1 is shut off and blocked at net heads above 7.80 m.
14-35 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-63E Solution We are to calculate the mass flow rate of slurry through a two-lobe rotary pump for given values of lobe volume and rotation rate. Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between lobes or between lobes and the casing. 3 The slurry is incompressible. Analysis By studying the figure provided with this problem, we see that for each 360o rotation of the two counterrotating shafts (n = 1 rotation), the total volume of pumped fluid is Vclosed = 4Vlobe
(1)
4 ( 0.145 gal ) V V = n closed = ( 300 rot/min ) = 174 gal/min 1 rot n
(2)
Closed volume pumped per rotation: The volume flow rate is then calculated from Eq. 14-11,
Thus, the volume flow rate is 174 gpm. Discussion If there were leaks in the pump, the volume flow rate would be lower. The fluid density was not used in the calculation of volume flow rate. However, the higher the density, the higher the required shaft torque and brake horsepower.
14-64E Solution We are to calculate the mass flow rate of slurry through a three-lobe rotary pump for given values of lobe volume and rotation rate. Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between lobes or between lobes and the casing. 3 The slurry is incompressible. Analysis When there are three lobes, three lobe volumes are pumped for each 360o rotation of each rotor (n = 1 rotation). Thus, the total volume of pumped fluid is Vclosed = 6Vlobe
(1)
6 ( 0.087 gal ) V V = n closed = ( 300 rot/min ) = 157 gal/min 1 rot n
(2)
Closed volume pumped per rotation: The volume flow rate is then calculated from Eq. 14-11,
Thus, the volume flow rate is 157 gpm. Discussion If there were leaks in the pump, the volume flow rate would be lower. This flow rate is slightly lower than that of the previous problem. Why? For the same overall diameter, it is clear from geometry that the more lobes, the less the volume per lobe, and the more “wasted” volume inside the pump.
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Chapter 14 Turbomachinery 14-65 Solution We are to calculate the volume flow rate of tomato paste through a positive-displacement pump for given values of lobe volume and rotation rate. Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between lobes or between lobes and the casing. 3 The fluid is incompressible. Analysis By studying Fig. 14-27 or 14-30, we see that for each 360o rotation of the two counter-rotating shafts (n = 1 rotation), the total volume of pumped fluid is
Closed volume pumped per rotation:
Vclosed = 4Vlobe
(1)
) = 5820 cm /min
(2)
The volume flow rate is then calculated from Eq. 14-11, V V = n closed = ( 400 rot/min ) n
(
4 3.64 cm3 1 rot
3
Thus, the volume flow rate is 5820 cm3/min. Discussion If there were leaks in the pump, the volume flow rate would be lower. The fluid density was not used in the calculation of volume flow rate. However, the higher the fluid density, the higher the required shaft torque and brake horsepower.
14-66 Solution
We are to calculate the volume flow rate per rotation of a gear pump.
Assumptions 1 The flow is steady in the mean. 2 There are no leaks in the gaps between lobes or between lobes and the casing. 3 The fluid is incompressible. Analysis From Fig. 14-26c, we count 14 teeth per gear. Thus, for each 360o rotation of each gear (n = 1 rotation), 3 14⋅(0.350 cm ) of fluid is pumped. Sine there are two gears, the total volume of fluid pumped per rotation is 2(14)(0.350 cm3) = 9.80 cm3. Discussion
The actual volume flow rate will be lower than this due to leakage in the gaps.
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Chapter 14 Turbomachinery 14-67 Solution We are to calculate the brake horsepower and net head of an idealized centrifugal pump at a given volume flow rate and rotation rate. Assumptions 1 The flow is steady in the mean. 2 The water is incompressible. 3 The efficiency of the pump is 100% (no irreversible losses). Properties
We take the density of water to be ρ = 998.0 kg/m3.
Analysis Since the volume flow rate (capacity) is given, we calculate the normal velocity components at the inlet and the outlet using Eq. 14-12, V1,n =
Normal velocity, inlet:
V 0.573 m3 /s = = 4.222 m/s 2π r1b1 2π ( 0.120 m )( 0.180 m )
V1 = V1,n, and V1,t = 0, since α1 = 0o. Similarly, V2,n = 2.714 m/s and Tangential component of absolute velocity at impeller outlet: V2,t = V2,n tan α 2 = ( 2.714 m/s ) tan(35o ) = 1.900 m/s Now we use Eq. 14-17 to predict the net head, H=
ω ⎛⎜
⎞ 78.54 rad/s r2V2,t − r1 V1,t ⎟ = ( 0.240 m )(1.900 m/s ) = 3.65 m N ⎟ 9.81 m/s 2 g⎜ 0 ⎠ ⎝
Finally, we use Eq. 14-16 to predict the required brake horsepower,
(
)(
bhp = ρ gV H = 998. kg/m3 9.81 m/s 2
⎛
)( 0.573 m /s ) ( 3.65 m ) ⎜ kgW⋅⋅ms 3
⎝
3 2
⎞ ⎟ = 20, 500 W ⎠
Discussion The actual net head delivered to the water will be lower than this due to inefficiencies. Similarly, actual brake horsepower will be higher than that predicted here due to inefficiencies in the pump, friction on the shaft, etc.
14-38 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-68 Solution For given flow conditions and stator blade shape at a given radius, we are to design the rotor blade. Specifically, we are to calculate the leading and trailing edge angles of the rotor blade and sketch its shape. We are also to decide how many rotor blades to construct.
Stator blade row G Vin
Assumptions 1 The air is nearly incompressible. 2 The flow area between hub and tip is constant. 3 Two-dimensional blade row analysis is appropriate.
G Vin G Vst
26.6o
Analysis First we analyze flow through the stator from an absolute reference frame, using the two-dimensional approximation of a cascade (blade row) of stator blades as sketched in Fig. 1. Flow enters axially (horizontally), and is turned 26.6o downward. Since the axial component of velocity must remain constant to conserve mass, the magnitude of the G velocity leaving the trailing edge of the stator, Vst is calculated,
FIGURE 1 Analysis of the stator of a vaneaxial flow fan as a two-dimensional cascade of stator blades; absolute reference frame.
Rotor blade row
ωr
G Vst
ωr
βrt
G Vst, relative
Direction of G Vrt, relative
G G Vrt = Vin
G Vrt, relative
βrt βrl
FIGURE 2 Analysis of the stator trailing edge velocity as it impinges on the rotor leading edge; relative reference frame.
FIGURE 3 Analysis of the rotor trailing edge velocity; absolute reference frame.
Vst =
Vin 31.4 m/s = = 35.12 m/s cosβ st cos 26.6o
(
)
(1)
G The direction of Vst is assumed to be that of the rotor trailing edge. In other words we assume that the flow turns nicely through the blade row and exits parallel to the trailing edge of the blade, as shown in the sketch. G We convert Vst to the relative reference frame moving with the rotor blades. At a radius of 0.50 m, the tangential velocity of the rotor blades is
14-39 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery ⎡ ⎛ 2π rad ⎞⎛ 1 min ⎞ ⎤ uθ = ω r = ⎢(1800 rot/min ) ⎜ ⎟⎜ ⎟ ⎥ ( 0.50 m ) = 94.25 m/s ⎝ rot ⎠⎝ 60 s ⎠ ⎦ ⎣
(2)
Since the rotor blade row moves upward in the figure provided with this problem, we add a downward velocity with G magnitude given by Eq. 2 to translate Vst into the rotating reference frame sketched in Fig. 2. The angle of the leading edge of the rotor, βrl, can be calculated. After some trig,
β rl = arctan
ω r + Vin tan β st Vin
= arctan
(
94.25 m/s + ( 31.4 m/s ) tan 26.6o 31.4 m/s
) = 74.06
o
(3)
The air must now be turned by the rotor blade row in such a way that it leaves the trailing edge of the rotor blade at zero angle (axially – no swirl) from an absolute reference frame. This determines the rotor’s trailing edge angle, βrt. Specifically, when we add an upward velocity of magnitude ωr (Eq. 2) to the relative velocity exiting the trailing edge of G G the rotor, Vrt, relative , we convert back to the absolute reference frame, and obtain Vrt , the velocity leaving the rotor trailing G G G edge. It is this velocity, Vrt , which must be axial (horizontal). Furthermore, to conserve mass Vrt must equal Vin since we G are assuming incompressible flow. Working “backwards” we construct Vrt, relative in Fig. 3. Some trigonometry reveals that
β rt = arctan
ωr Vin
= arctan
94.25 m/s = 71.57o 31.4 m/s
(4)
We conclude that the rotor blade at this radius has a leading edge angle of about 74.1o (Eq. 3) and a trailing edge angle of about 71.6o (Eq. 4). A sketch of the rotor blade is provided in Fig. 2; it is clear that the blade is nearly straight, at least at this radius. Finally, to avoid interaction of the stator blade wakes with the rotor blade leading edges, we choose the number of rotor blades such that it has no common denominator with the number of stator blades. Since there are 18 stator blades, we pick a number like 13, 17, or 19 rotor blades. 16 would not be appropriate since it shares a common denominator of 2 with the number 18. Discussion
We can easily repeat the calculation for all radii from hub to tip, completing the design of the entire rotor.
14-40 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
Turbines
14-69C Solution
We are to discuss why turbines have higher efficiencies than pumps.
Analysis There are several reasons for this. First, pumps normally operate at higher rotational speeds than do turbines; therefore, shear stresses and frictional losses are higher. Second, conversion of kinetic energy into flow energy (pumps) has inherently higher losses than does the reverse (turbines). You can think of it this way: Since pressure rises across a pump (adverse pressure gradient), but drops across a turbine (favorable pressure gradient), boundary layers are less likely to separate in a turbine than in a pump. Third, turbines (especially hydroturbines) are often much larger than pumps, and viscous losses become less important as size increases. Finally, while pumps often operate over a wide range of flow rates, most electricity-generating turbines run within a narrower operating range and at a controlled constant speed; they can therefore be designed to operate very efficiently at those conditions. Discussion
Students’ answers should be in their own words.
14-70C Solution
We are to discuss the classification of dynamic pumps and reaction turbines.
Analysis Dynamic pumps are classified according to the angle at which the flow exits the impeller blade – centrifugal, mixed-flow, or axial. Reaction turbines, on the other hand, are classified according to the angle that the flow enters the runner – radial, mixed-flow, or axial. This is the main difference between how dynamic pumps and reaction turbines are classified. Discussion
Students’ answers should be in their own words.
14-71C Solution
We are to discuss the meaning and purpose of draft tubes.
Analysis A draft tube is a diffuser that also turns the flow downstream of a turbine. Its purpose is to turn the flow horizontally and recover some of the kinetic energy leaving the turbine runner. If the draft tube is not designed carefully, much of the kinetic energy leaving the runner would be wasted, reducing the overall efficiency of the turbine system. Discussion
Students’ answers should be in their own words.
14-72C Solution
We are to name and describe the two types of dynamic turbine.
Analysis There are two basic types of dynamic turbine – impulse and reaction. In an impulse turbine, fluid is sent through a nozzle so that most of its available mechanical energy is converted into kinetic energy. The high-speed jet then impinges on bucket-shaped vanes that transfer energy to the turbine shaft. In a reaction turbine, the fluid completely fills the casing, and the runner is rotated by momentum exchange due to pressure differences across the blades, rather than by kinetic energy impingement. Impulse turbines require a higher head, but can operate with a smaller volume flow rate. Reaction turbines can operate with much less head, but require higher volume flow rate. Discussion
Students’ answers should be in their own words.
14-41 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-73C Solution
We are to discuss reverse swirl in reaction turbines.
Analysis Reverse swirl is when the runner blades turn the flow so much that the swirl at the runner outlet is in the direction opposite to runner rotation. Reverse swirl is desirable so that more power is absorbed from the water. We can easily see this from the Euler turbomachine equation,
Wshaft = ω Tshaft = ρωV ( r2V2,t − rV 1 1,t )
(1)
Namely, since there is a negative sign on the last term, the shaft power increases if V1,t is negative, in other words, if there is reverse swirl at the runner outlet. If there is too much reverse swirl, a lot of extra kinetic energy gets wasted downstream of the runner. Discussion A well designed draft tube can recover a good portion of the streamwise kinetic energy of the water leaving the runner. However, the swirling kinetic energy cannot be recovered.
14-74 Solution We are to prove that the maximum power of a Pelton wheel occurs when ωr = Vj/2 (bucket moving at half the jet speed). Assumptions 1 Frictional losses are negligible, so that the Euler turbomachine equation applies, and the relative exit speed of the jet is the same as its relative inlet speed. 2 The turning angle is sufficient to prevent the exiting fluid from striking the next bucket. 3 The fluid, jet speed, volume flow rate, turning angle, and wheel radius are fixed – rotation rate is the only variable about which we are concerned. Analysis
With the stated assumptions, Eq. 14-40 applies, namely,
(
)
Wshaft = ρω rV V j − ω r (1 − cos β )
Output shaft power:
(1)
We differentiate Eq. 1 with respect to ω, and set the derivative equal to zero, Maximum power: dWshaft =0 dω
→
(
)
d ωV j − ω 2 r = 0 dω
→
V j − 2ω r = 0
(2)
Solution of Eq. 2 yields the desired result, namely, the maximum power of a Pelton wheel occurs when ωr = Vj/2 (bucket moving at half the jet speed). Discussion To be sure that we have not identified a minimum instead of a maximum, we could substitute some numerical values and plot Wshaft versus ω. It turns out that we have indeed found the maximum value of Wshaft .
14-42 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-75 Solution
We are to calculate several performance parameters for a Pelton wheel turbine.
Assumptions 1 Frictional losses are negligible, so that the Euler turbomachine equation applies, and the relative exit speed of the jet is the same as its relative inlet speed. 2 The turning angle is sufficient to prevent the exiting fluid from striking the next bucket. 3 The water is at 20oC. Properties
The density of water at T = 20oC is 998.0 kg/m3.
Analysis
(a) The volume flow rate of the jet is equal to jet area times jet velocity, 2 V = V j π D j 2 / 4 = (102 m/s ) π ( 0.100 m ) / 4 = 0.801 m 3 /s
(b) The maximum output shaft power occurs when the bucket moves at half the jet speed (ωr = Vj/2). Thus,
ω=
Vj 2r
→
=
102 m/s = 27.87 rad/s 2 (1.83 m )
⎛ rot ⎞ ⎛ 60 s ⎞ n = ( 27.87 rad/s ) ⎜ ⎟⎜ ⎟ = 266 rpm ⎝ 2π rad ⎠ ⎝ min ⎠
(c) The ideal shaft power is found from Eq. 14-40,
(
)
Wideal = ρω rV V j − ω r (1 − cos β ) ⎛ rad ⎞ m3 ⎛ = 998.0 kg/m3 ⎜ 27.87 (1.83 m ) ⎜ 0.801 ⎟ s ⎠ s ⎝ ⎝
(
)
⎞ ⎛ 102 m/s ⎞ o ⎟⎜ ⎟ 1 − cos165 2 ⎠ ⎠⎝
(
)
⎛ N ⋅ s 2 ⎞ ⎛ W ⋅ s ⎞ ⎛ 1 MW ⎞ ×⎜ ⎟⎜ ⎟⎜ 6 ⎟ = 4.09 MW ⎝ kg ⋅ m ⎠ ⎝ N ⋅ m ⎠ ⎝ 10 W ⎠
where we have substituted ωr = Vj/2 for convenience in the calculations. Since the turbine efficiency is given, we calculate the actual output shaft power, or brake horsepower, Wactual = bhp = Widealηturbine = ( 4.09 MW )( 0.82 ) = 3.35 MW
Discussion
At other rotation speeds, the turbine would not be as efficient.
14-76 Solution
We are to estimate the ideal power production from a hydroturbine.
Assumptions
1 Frictional losses are negligible. 2 The water is at 20oC.
Properties
The density of water at T = 20oC is 998.0 kg/m3.
Analysis
The ideal power produced by a hydroturbine is ⎛ N ⋅ s 2 ⎞ ⎛ W ⋅ s ⎞ ⎛ 1 MW ⎞ Wideal = ρ gV H gross = 998.0 kg/m3 9.81 m/s 2 1.5 m3 /s ( 650 m ) × ⎜ ⎟⎜ ⎟⎜ 6 ⎟ = 9.55 MW ⎝ kg ⋅ m ⎠ ⎝ N ⋅ m ⎠ ⎝ 10 W ⎠
(
)(
)(
)
Discussion If a hydroelectric dam were to be built on this site, the actual power output per turbine would be smaller than this, of course, due to inefficiencies.
14-43 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-77E Solution
We are to estimate the power production from a hydroelectric plant.
Properties
The density of water at T = 70oF is 62.30 lbm/ft3.
Analysis
The ideal power produced by one hydroturbine is
Wideal = ρ gV H gross
(
)(
= 62.30 lbm/ft 3 32.2 ft/s 2
⎛
lbf ⋅ s 2 ⎞ ⎛ ft 3 ⎞ ⎛ 1.356 W ⎞ ⎛ 1 min ⎞ ⎛ 1 MW ⎞ 0.1337 ⎟⎜ ⎟ gal ⎠ ⎜⎝ ft ⋅ lbf/s ⎟⎠ ⎜⎝ 60 s ⎟⎠ ⎜⎝ 106 W ⎟⎠ ⎝ 32.2 lbm ⋅ ft ⎠ ⎝
) ( 203, 000 gal/min )(1065 ft ) × ⎜
= 40.70 MW
But inefficiencies in the turbine, the generator, and the rest of the system reduce the actual electrical power output. For each turbine, Welectrical = Widealη turbineηgeneratorηother = ( 40.70 MW )( 0.952 )( 0.945 )(1 − 0.035 ) = 35.3 MW
Finally, since there are 12 turbines in parallel, the total power produced is W total, electrical = 12Welectrical = 12 ( 35.3 MW ) = 424 MW
Discussion A small improvement in any of the efficiencies ends up increasing the power output, and increases the power company’s profitability.
14-44 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-78 Solution
We are to calculate runner blade angles, required net head, and power output for a Francis turbine.
Assumptions 1 The flow is steady. 2 The fluid is water at 20oC. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties
For water at 20oC, ρ = 998.0 kg/m3.
Analysis
We solve for the normal component of velocity at the inlet, V2,n =
V 340 m3 /s = = 37.0 m/s 2π r2 b2 2π ( 2.00 m )( 0.731 m )
(1)
Using the figure provided with this problem as a guide, the tangential velocity component at the inlet is V2,t = V2,n tan α 2 = ( 37.0 m/s ) tan 30o = 21.4 m/s
(2)
The angular velocity is ω = 2π n /60 = 18.85 rad/s. The tangential velocity component of the absolute velocity at the inlet is obtained from trigonometry to be (see Eq. 14-45) V2,t = ω r2 −
V2,n tan β 2
From the above relationship, we solve for the runner leading edge angle β2, ⎡ ⎤ ⎤ 37.0 m/s o ⎥ = 66.2 ⎥ = arctan ⎢ − − ω r V 18.85 rad/s 2.00 m 21.4 m/s )( ) ⎥ 2,t ⎦ ⎣⎢ 2 ⎣⎢ ( ⎦⎥ ⎡
β 2 = arctan ⎢
V2,n
(3)
Equations 1 through 3 are repeated for the runner outlet, with the following results: V1,n = 17.3 m/s,
Runner outlet:
V1,t = 3.05 m/s,
β1 = 36.1o
(4)
Using Eqs. 2 and 4, the shaft output power is estimated from the Euler turbomachine equation,
(
)
(
)
3 Wshaft = ρωV ( r2V2,t − rV (18.85 rad/s ) 340 m3 /s × ⎡⎣( 2.00 m )( 21.4 m/s ) − (1.42 m )( 3.05 m/s )⎤⎦ 1 1,t ) = 998.0 kg/m
= 2.46 × 108
kg ⋅ m ⎛ MW ⋅ s ⎜ s3 ⎝ 106 kg ⋅ m 2 2
3
⎞ ⎟ = 246.055 MW ≅ 246 MW ⎠
(5)
Finally, we calculate the required net head using Eq. 14-44, assuming that ηturbine = 100% since we are ignoring irreversibilities, H=
⎛ 106 kg ⋅ m 2 bhp 246.055 MW = ⎜ ρ gV ( 998.0 kg/m3 )( 9.81 m/s 2 )( 340 m3 /s ) ⎝ MW ⋅ s3
⎞ ⎟ = 73.9 m ⎠
(6)
Since the required net head is less than the gross net head, the design is feasible. Discussion This is a preliminary design in which we are neglecting irreversibilities. Actual output power will be lower, and actual required net head will be higher than the values predicted here.
14-45 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-79 [Also solved using EES on enclosed DVD] Solution We are to examine the effect of runner outlet angle α1 on the required net head and the output power of a hydroturbine. Assumptions 1 The flow is steady. 2 The fluid is water at 20oC. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties
For water at 20 C, ρ = 998.0 kg/m . o
100
700
Hgross
600
80
H
500
3
60
Analysis We repeat the calculations of the previous problem, but with α1 varying from -20o to 20o. The results are plotted. We see that the required net head H and the output brake horsepower bhp decrease as α1 increases. This agrees with our expectations, based on the negative sign on the V1,t term in the Euler turbomachine equation. In other words, we can produce greater power by increasing the reverse swirl, but at the cost of increased required net head. However, when α1 is smaller than about -9o, the required net head rises above Hgross, which is impossible, even with no irreversibilities. Thus, when α1 < -9o, the predicted net head and brake horsepower are not feasible – they violate the second law of thermodynamics. Discussion
bhp (MW)
H (m)
400 300
40
bhp
200
20
100 0
0 -20
-10
0
10
20
α1 (degrees)
A small amount of reverse swirl is usually good, but too much reverse swirl is not good.
14-46 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-80 Solution
We are to calculate flow angles, required net head, and power output for a Francis turbine.
Assumptions 1 The flow is steady. 2 The fluid is water at 20oC. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties
For water at 20oC, ρ = 998.0 kg/m3.
Analysis inlet,
The angular velocity is ω = 2π n /60 = 10.47 rad/s. We solve for the normal component of velocity at the
V2,n =
80.0 m3 /s V = = 7.490 m/s 2π r2 b2 2π ( 2.00 m )( 0.85 m )
(1)
The tangential velocity component of the absolute velocity at the inlet is obtained from trigonometry to be (see Eq. 14-45) V2,t = ω r2 −
V2,n tan β 2
= (10.47 rad/s )( 2.00 m ) −
7.490 m/s = 17.61 m/s tan 66o
(2)
From these two components of V2 in the absolute coordinate system, we calculate the angle α2 through which the wicket gates should turn the flow,
α 2 = tan −1
V2,t V2,n
= tan −1
17.61 m/s = 66.95o ≈ 67o 7.490 m/s
(3)
In exactly similar fashion, we solve for the velocity components and swirl angle at the runner outlet. We get Runner outlet: V1,n = 4.664 m/s,
V1,t = −0.3253 m/s,
(4)
α1 = -3.990o ≈ -4.0o
Since α1 is negative, this turbine operates with a small amount of reverse swirl. Using the tangential velocity components of Eqs. 2 and 4, the shaft output power is estimated from the Euler turbomachine equation,
(
)
(
3 Wshaft = ρωV ( r2V2,t − rV (10.47 rad/s ) 80.0 m3 /s 1 1,t ) = 998.0 kg/m
× ⎡⎣( 2.00 m )(17.61 m/s ) − (1.30 m )( −0.3253 m/s ) ⎤⎦ kg ⋅ m 2 = 2.98 × 10 s3 7
⎛ MW ⋅ s3 ⎜ 6 2 ⎝ 10 kg ⋅ m
⎞ ⎟ = 29.8 MW ⎠
) (5)
Finally, we calculate the required net head using Eq. 14-44, assuming that ηturbine = 100% since we are ignoring irreversibilities, H=
⎛ 106 kg ⋅ m 2 ⎞ bhp 29.8 MW = ⎜ ⎟ = 38.0 m ρ gV ( 998.0 kg/m3 )( 9.81 m/s 2 )( 80.0 m3 /s ) ⎝ MW ⋅ s3 ⎠
(6)
Discussion This is a preliminary design in which we are neglecting irreversibilities. Actual output power will be lower, and actual required net head will be higher than the values predicted here. Notice the double negative in the outlet terms of Eq. 5 – reverse swirl leads to greater performance, but requires more head.
14-47 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-81E Solution
We are to calculate flow angles, required net head, and power output for a Francis turbine.
Assumptions 1 The flow is steady. 2 The fluid is water at 68oF. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties
For water at 68oF, ρ = 62.32 lbm/ft3.
Analysis The angular velocity is ω = 2π n /60 = 12.57 rad/s. The volume flow rate is (4.70 × 106 gpm)[ft3/s / (448.83 gpm)] = 10,470 ft3/s. We solve for the normal component of velocity at the inlet, V2,n =
10, 470 ft 3 /s V = = 97.1 ft/s 2π r2 b2 2π ( 6.60 ft )( 2.60 ft )
(1)
The tangential velocity component of the absolute velocity at the inlet is obtained from trigonometry to be (see Eq. 14-45) V2,t = ω r2 −
V2,n tan β 2
= (12.57 rad/s )( 6.60 ft ) −
97.12 ft/s = 69.3 ft/s tan 82o
(2)
From these two components of V2 in the absolute coordinate system, we calculate the angle α2 through which the wicket gates should turn the flow,
α 2 = tan −1
V2,t V2,n
= tan −1
69.3 ft/s = 35.5o 97.1 ft/s
(3)
In exactly similar fashion, we solve for the velocity components and swirl angle at the runner outlet. We get V1,n = 52.6 ft/s,
Runner outlet:
V1,t = 4.49 ft/s,
α1 = 4.9o
(4)
Since α1 is positive, this turbine operates with a small amount of forward swirl. Using the tangential velocity components of Eqs. 2 and 4, the shaft output power is estimated from the Euler turbomachine equation,
(
)
(
3 Wshaft = ρωV ( r2V2,t − rV (12.57 rad/s ) 10, 470 ft 3 /s 1 1,t ) = 62.32 lbm/ft
)
⎛ ⎞ lbf ⋅ s × ⎡⎣( 6.60 ft )( 69.3 ft/s ) − ( 4.40 ft )( 4.49 ft/s ) ⎤⎦ ⎜ ⎟ ⎝ 32.174 lbm ⋅ ft ⎠ −3 ft ⋅ lbf ⎛ 1.8182 × 10 hp ⋅ s ⎞ 5 = 1.116 × 108 ⎜ ⎟ = 2.03 × 10 hp s ⎝ ft ⋅ lbf ⎠ 2
(5)
Finally, we calculate the required net head using Eq. 14-44, assuming that ηturbine = 100% since we are ignoring irreversibilities, H=
bhp 1.116 × 108 ft ⋅ lbf/s ⎛ 32.174 lbm ⋅ ft ⎞ = ⎜ ⎟ = 171 ft 3 2 3 lbf ⋅ s 2 ρ gV ( 62.32 lbm/ft )( 32.174 ft/s )(10, 470 ft /s ) ⎝ ⎠
(6)
Discussion This is a preliminary design in which we are neglecting irreversibilities. Actual output power will be lower, and actual required net head will be higher than the values predicted here.
14-48 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-82E Solution We are to calculate the runner blade trailing edge angle such that there is no swirl. At this value of β1, we are to also calculate the shaft power. Assumptions 1 The flow is steady. 2 The fluid is water at 68oF. 3 The blades are infinitesimally thin. 4 The flow is everywhere tangent to the runner blades. 5 We neglect irreversible losses through the turbine. Properties
For water at 68oF, ρ = 62.32 lbm/ft3.
Analysis Using EES or other software, we adjust β1 by trial and error until α1 = 0. It turns out that β1 = 43.6o, at which Wshaft = 2.12 × 105 hp. Discussion It turns out that the swirl angle at the runner output is a strong function of β1 – a small change in β1 leads to a large change in α1. The shaft power increases, as expected, since the original swirl angle was positive. The increase in shaft power is less than 5%.
Pump and Turbine Scaling Laws
14-83C Solution
We are to give a definition of “affinity”, and explain why the scaling laws are called “affinity laws”.
Analysis Among the many definitions of “affinity” is “inherent likeness or agreement”, and “...resemblance in general plan or structure”. When two pumps or two turbines are geometrically similar and operate under dynamically similar conditions, they indeed have “inherent likeness”. Thus, the phrase “affinity laws” is appropriate for the scaling laws of turbomachinery. Discussion
Students will have various definitions, depending on the dictionary they use.
14-84C Solution (a) True: Rotation rate appears with an exponent of 1 in the affinity law for capacity. Thus, the change is linear. (b) False: Rotation rate appears with an exponent of 2 in the affinity law for net head. Thus, if the rpm is doubled, the net head increases by a factor of 4. (c) False: Rotation rate appears with an exponent of 3 in the affinity law for shaft power. Thus, if the rpm is doubled, the shaft power increases by a factor of 8. (d) True: The affinity laws apply to turbines as well as pumps, so this statement is true, as discussed in Part (c).
14-85C Solution We are to discuss which pump and turbine performance parameters are used as the independent parameter, and explain why Analysis For pumps, we use CQ, the capacity coefficient, as the independent parameter. The reason is that the goal of a pump is to move fluid from one place to another, and the most important parameter is the pump’s capacity (volume flow rate). On the other hand, for most turbines, we use CP, the power coefficient, as the independent parameter. The reason is that the goal of a turbine is to rotate a shaft, and the most important parameter is the turbine’s brake horsepower Discussion There are exceptions. For example, when analyzing a positive-displacement turbine used to measure volume flow rate, capacity is more important than output shaft power, so one might use CQ, instead of CP as the independent parameter.
14-49 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-86 Solution curve.
We are to nondimensionalize a pump performance 4
Analysis expression
The pump’s net head is approximated by the parabolic H = H 0 − aV 2
Pump performance curve:
3
(1)
where shutoff head H0 = 24.4 m of water column, coefficient a = 0.0678 m/Lpm2, available pump head Havailable is in units of meters of water column, and capacity V is in units of liters per minute (Lpm). By definition, head coefficient CH = (gH)/(ω2D2), and capacity coefficient CQ = ( V )/(ωD3). To convert, we must be careful with units. For example, we must convert the rotation rate from rpm to radians per second. The rotational speed is
ω = 4200
CH 2 1
0 0
rot ⎛ 2π rad ⎞ ⎛ 1 min ⎞ rad = 439.823 min ⎜⎝ rot ⎟⎠ ⎜⎝ 60 s ⎟⎠ s
0.05
0.1
0.15
CQ
Sample calculations at 14.0 Lpm are shown below. CQ =
14.0 L/min V = ω D 3 ( 439.823 rad/s )( 0.0180 m )3
⎛ 1 m3 ⎞ ⎛ 1 min ⎞ ⎜ ⎟⎜ ⎟ = 0.90966 ≅ 0.0910 ⎝ 1000 L ⎠ ⎝ 60 s ⎠
At this flow rate (14.0 Lpm), the net head is obtained from Eq. 1,
(
)
H = 24.4 m − 0.0678 m/Lpm 2 (14.0 Lpm ) = 11.111 m ≅ 11.1 m 2
from which the head coefficient can be calculated, CH =
(
)
9.81 m/s 2 (11.111 m ) gH = = 1.7391 ≅ 1.74 ω 2 D 2 ( 439.823 rad/s )2 ( 0.0180 m )2
These calculations are repeated for a range of volume flow rates from 0 to Vmax . The nondimensionalized pump performance curve is plotted. Discussion Since radians is a dimensionless unit, the units of CH and CQ are unity. In other words, CH and CQ are nondimensional parameters, as they must be.
14-50 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-87 Solution
We are to calculate the specific speed of a water pump, and determine what kind of pump it is.
Analysis First, we use the values of CH and CQ calculated in the previous problem at the BEP to calculate the dimensionless form of NSp,
Dimensionless pump specific speed:
N Sp =
CQ1/ 2 CH 3 / 4
=
0.0909661/ 2 = 0.1992 1.73913 / 4
Thus, NSp = 0.199, and is dimensionless. From the conversion given in the text, NSp,US = 0.1992 × 2,734 = 545. Alternatively, we can use the original dimensional data to calculate NSp,US, Dimensional pump specific speed in customary US units:
( n, rpm ) (V , gpm ) 3/ 4 ( H , ft )
⎛
1/ 2
N Sp ,US =
=
⎛ 0.2642 gpm ⎞ ⎟ Lpm ⎝ ⎠
( 4200 rpm ) ⎜⎜14.0 Lpm ⎜ ⎝
⎛ ft ⎛ ⎞⎞ ⎜11.111 m ⎜ ⎟⎟ 0.3048 m ⎝ ⎠⎠ ⎝
3/ 4
1/ 2
⎞ ⎟⎟ ⎠
= 545
From either the dimensionless or the dimensional pump specific speed, Fig. 14-73 shows that this is definitely a centrifugal pump. Discussion We calculated the dimensional pump specific speed two ways as a check of our algebra, and (fortunately) the results agree.
14-51 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-88 Solution Analysis expression
We are to nondimensionalize a fan performance curve. The fan’s net head is approximated by the parabolic 2
Fan performance curve:
H = H 0 − aV 2
(1)
where shutoff head H0 = 60.0 mm of water column, coefficient a = 2.50 × 10-7 mm/Lpm2, available fan head Havailable is in units of mm of water column, and capacity V is in units of liters per minute (Lpm). By definition, head coefficient CH = (gH)/(ω2D2), and capacity coefficient CQ = ( V )/(ωD3). To convert, we must be careful with units. For example, we must convert the rotation rate from rpm to radians per second. The rotational speed becomes
ω = 600
CH 1
rot ⎛ 2π rad ⎞ ⎛ 1 min ⎞ rad = 62.8319 ⎜ ⎟ ⎜ ⎟ min ⎝ rot ⎠ ⎝ 60 s ⎠ s
0 0
0.05
0.1 CQ
0.15
0.2
Sample calculations at 13,600 Lpm are shown below. Capacity coefficient at 13,600 Lpm:
CQ =
13, 600 L/min V = ω D 3 ( 62.8319 rad/s )( 0.300 m )3
⎛ 1 m3 ⎞ ⎛ 1 min ⎞ ⎜ ⎟⎜ ⎟ = 0.13361 ≅ 0.134 ⎝ 1000 L ⎠ ⎝ 60 s ⎠
At this flow rate, the net head is obtained from Eq. 1,
(
)
H = 60.0 mm − 2.50 × 10 −7 mm/Lpm 2 (13, 600 Lpm ) = 13.76 mm H 2 O 2
from which the head coefficient can be calculated, Head coefficient at 13,600 Lpm:
(
)
9.81 m/s 2 ( 0.01376 m H 2 O ) 998. kg/m3 ( m air ) gH = 0.32023 ≅ 0.320 CH = 2 2 = 2 2 3 ω D ( 62.8319 rad/s ) ( 0.300 m ) 1.184 kg/m ( m H 2 O )
Note the ratio of water density to air density to convert the head from water column height to air column height. The calculations are repeated for a range of volume flow rates from 0 to Vmax . The nondimensionalized pump performance curve is plotted. Discussion Since radians is a dimensionless unit, the units of CH and CQ are unity. In other words, CH and CQ are nondimensional parameters, as they must be.
14-52 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-89 Solution
We are to calculate the specific speed of a fan, and determine what kind of fan it is.
Analysis First, we use the values of CH and CQ calculated in the previous problem at the BEP (to four significant digits) to calculate the dimensionless form of NSp, N Sp =
Dimensionless pump specific speed:
CQ1/ 2 CH 3 / 4
=
0.133611/ 2 = 0.8587 0.320233 / 4
Thus, NSp = 0.859, and is dimensionless. From the conversion given in the text, NSp,US = 0.8587 × 2,734 = 2350. Alternatively, we can use the dimensional data to calculate NSp,US, the dimensional pump specific speed in US units,
( n, rpm ) (V , gpm ) N Sp ,US = 3/ 4 ( H , ft )
1/ 2
=
⎛ ⎛ 0.2642 gpm ⎞ ( 600 rpm ) ⎜⎜13, 600 Lpm ⎜ ⎟ Lpm ⎝ ⎠ ⎝ ⎛ ft ⎛ ⎞⎞ ⎜11.60 m ⎜ ⎟⎟ 0.3048 m ⎝ ⎠⎠ ⎝
1/ 2
⎞ ⎟⎟ ⎠
= 2350
3/ 4
Note that we use H = 11.60 m of air, since air is the fluid being pumped here. We calculate H as Hair = Hwater(ρwater/ρair) = (0.01376 m of water)(998/1.184) = 11.60 m of air. From either the dimensionless or the dimensional pump specific speed, Fig. 14-73 shows that this is most likely a centrifugal fan (probably a squirrel cage fan). Discussion We calculated the dimensional pump specific speed two ways as a check of our algebra, and (fortunately) the results agree.
14-90 Solution
We are to calculate the specific speed of a water pump, and determine what kind of pump it is.
Analysis form of NSp,
First, we use the values of CH and CQ calculated in Example 14-11 at the BEP to calculate the dimensionless 1/ 2
N Sp =
Dimensionless pump specific speed:
CQ CH
3/ 4
=
0.01121/ 2 = 0.481 0.1333 / 4
Thus, NSp = 0.481. From the conversion given in the text, NSp,US = 1310. From Fig. 14-73, this is most likely a centrifugal pump. Discussion From Fig. 14-73, the maximum efficiency one can expect from a centrifugal pump at this value of NSp is about 0.87 (87%). The actual pump efficiency is only 81.2% (from Example 14-11), so there is room for improvement in the design.
14-91 Solution
We are to decide what king of pump should be designed for given performance criteria.
Analysis
We calculate the pump specific speed at the given conditions,
N Sp =
ωV 1/ 2
( gH )
3/ 4
⎛ 1 m3 ⎞ ⎛ 1 min ⎞ ⎛ 2π rad/s ⎞ ⎛ 18.0 L/min ⎜ ⎜ ⎟⎜ ⎟⎜ ⎟ ⎝ 60 rpm ⎠ ⎝ ⎝ 1000 L ⎠ ⎝ 60 s ⎠
(1200 rpm ) ⎜ =
((
)
9.81 m/s 2 (1.6 m )
)
3/ 4
1/ 2
⎞ ⎟⎟ ⎠
= 0.276
From Fig. 14-73, we see that Len should design a centrifugal pump. The maximum pump efficiency is about 0.75 (75%) (based on Fig. 14-73 again). Discussion This value of NSp is, in fact, on the low end of the curve for centrifugal pumps, so a centrifugal pump is the best Len can do, in spite of the low efficiency.
14-53 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-92E Solution
We are to decide what king of pump should be designed for given performance criteria.
Properties
The density of water at 77oF is 62.24 lbm/ft3.
Analysis
We calculate the pump specific speed at the given conditions,
N Sp =
ωV 1/ 2
( gH )
3/ 4
⎛ 0.13368 ft 3 ⎛ 2π rad/s ⎞ ⎛ 2,500 gal/min ( 300 rpm ) ⎜ ⎜ ⎜ ⎟⎜ ⎝ 60 rpm ⎠ ⎝ ⎝ 1 gal = 3/ 4 32.174 ft/s 2 ( 45 ft )
((
)
)
⎞ ⎛ 1 min ⎞ ⎟⎜ ⎟ ⎠ ⎝ 60 s ⎠
1/ 2
⎞ ⎟⎟ ⎠
= 0.316
From Fig. 14-73, we choose a centrifugal pump. The maximum pump efficiency is about 0.82 (82%) (based on Fig. 1473 again). From the definition of pump efficiency, bhp = ηpumpρ V gH. Thus, the required brake horsepower is bhp =
3 2 3 ⎞ lbf ⋅ s 2 1 hp ρ gHV ( 62.24 lbm/ft )( 32.174 ft/s ) ( 45 ft ) ( 5.57 ft /s ) ⎛ ⎛ ⎞ = ⎜ ⎟ = 19, 025 ft ⋅ lbf/s ⎜ ⎟ = 34.6 hp 0.82 32.174 lbm ft 550 ft lbf/s ⋅ ⋅ η pump ⎝ ⎠ ⎝ ⎠
Thus, we expect that the pump will require about 35 hp to turn the shaft. Discussion Most large water pumps are of the centrifugal variety. This problem may also be solved in terms of the dimensional pump specific speed in customary US units: NSp,US = 864.
14-93 Solution
We are to calculate the performance of a modified pump.
Assumptions same.
1 The modified pump operates at the best efficiency point. 2 Pump diameter and fluid properties remain the
Analysis At homologous points, the affinity laws are used to estimate the operating conditions of the modified pump. We let the original pump be Pump A, and the modified pump be Pump B. We get 3
Volume flow rate:
ω ⎛D ⎞ 600 rpm VB = VA B ⎜ B ⎟ = (18. L/min ) = 9.0 L/min 1200 rpm ωA ⎝ DA ⎠
1
and 2
2
2
⎛ω ⎞ ⎛ D ⎞ ⎛ 600 rpm ⎞ H B = H A ⎜ B ⎟ ⎜ B ⎟ = (1.6 m ) ⎜ ⎟ = 0.40 m D 1200 rpm ⎠ ω ⎝ A ⎠ ⎝ A ⎠ ⎝
Net head:
1
The volume flow rate of the modified pump is 9.0 L/min; the net head is 0.40 m. The pump specific speed of the modified pump is
N Sp =
ωV 1/ 2
( gH )
3/ 4
⎛ 1 m3 ⎞ ⎛ 1 min ⎞ ⎛ 2π rad/s ⎞ ⎛ ( 600 rpm ) ⎜ ⎟⎜ ⎟ ⎜⎜ 9.0 L/min ⎜ ⎟ ⎝ 60 rpm ⎠ ⎝ ⎝ 1000 L ⎠ ⎝ 60 s ⎠ = 3/ 4 9.81 m/s 2 ( 0.40 m )
((
)
)
1/ 2
⎞ ⎟⎟ ⎠
= 0.276
Thus, NSp = 0.276, which is the same as that of the original pump. This is not surprising since the two pumps operate at homologous points. Discussion When the rpm is cut in half, all else being equal, the volume flow rate of the pump decreases by a factor of two, while the net head decreases by a factor of four. This agrees with the mnemonic of Fig. 14-74. The specific speed of the two pumps must match since they operate at homologous points.
14-54 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-94 Solution
We are to prove the relationship between turbine specific speed and pump specific speed.
Assumptions 1 The parameters such as net head, volume flow rate, diameter, etc. are defined in similar fashion for the pump and the turbine. Analysis
First we write the definitions of pump specific speed and turbine specific speed, N Sp =
Pump specific speed:
ωV 1/ 2
( gH )
3/ 4
(1)
and
ω ( bhp )
1/ 2
Turbine specific speed:
N St =
( ρ ) ( gH ) 1/ 2
5/ 4
(2)
After some rearrangement of Eq. 2,
ωV 1/ 2 ⎛ bhp ⎞ N St = ⎟ 3/ 4 ⎜ ( gH ) ⎜⎝ ρ gHV ⎟⎠
1/ 2
(3)
We recognize the first grouping of terms in Eq. 3 as NSp and the second grouping of terms as the square root of turbine efficiency ηturbine. Thus, Final relationship:
N St = N Sp η turbine
(4)
Discussion Since turbine efficiency is typically large (90 to 95% for large hydroturbines), pump specific speed and turbine specific speed are nearly equivalent. Note that Eq. 4 does not apply to a pump running backwards as a turbine or vice-versa. Such devices, called pump-turbines, are addressed in the next problem.
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Chapter 14 Turbomachinery 14-95 Solution We are to prove the relationship between turbine specific speed and pump specific speed for the case of a pump-turbine operating at the same volume flow rate and rotational speed when acting as a pump and as a turbine. Assumptions 1 The parameters such as net head, volume flow rate, diameter, etc. are defined in similar fashion for the pump and the turbine. Analysis
First we write the definitions of pump specific speed and turbine specific speed, N Sp =
Pump specific speed:
ωV 1/ 2
( gH )
3/ 4
(1)
pump
and
ω ( bhpturbine )
1/ 2
N St =
Turbine specific speed:
( ρ ) ( gH turbine ) 1/ 2
5/ 4
(2)
Note that we have added subscripts “pump” and “turbine” on net head and brake horsepower since Hpump ≠ Hturbine and bhppump ≠ bhpturbine. After some rearrangement of Eq. 2, N St =
ωV 1/ 2
( gH turbine )
3/ 4
⎛ bhpturbine ⎜⎜ ⎝ ρ gH turbineV
1/ 2
⎞ ⎟⎟ ⎠
=
ωV 1/ 2
(ηturbine )
1/ 2
( gH turbine )
3/ 4
(3)
We also write the definitions of pump efficiency and turbine efficiency,
ηpump =
ρ gV H pump
η turbine =
bhppump
bhpturbine
ρ gV H turbine
(4)
We solve both parts of Eq. 4 for V and equate the two, since V is the same whether the pump-turbine is operating as a pump or as a turbine. Eliminating V , H turbine =
H pump bhpturbine bhppumpηpumpηturbine
(5)
where ρ and g have also dropped out since they are the same for both cases. We plug Eq. 5 into Eq. 3 and rearrange, N St =
ωV 1/ 2
( gH )
η 3 / 4 ( turbine )
1/ 2
pump
⎛ bhppumpηpumpη turbine ⎞ ⎜ ⎟ bhpturbine ⎝ ⎠
3/ 4
(6)
We recognize the first grouping of terms in Eq. 6 as NSp and rearrange, Final relationship:
N St = N Sp (ηturbine )
5/ 4
(η ) pump
3/ 4
⎛ bhppump ⎞ ⎜ ⎟ ⎝ bhpturbine ⎠
3/ 4
(7)
An alternate expression is obtained in terms of net heads by substitution of Eq. 5, Alternate final relationship:
N St = N Sp ηturbine
⎛ H pump ⎞ ⎜ ⎟ ⎝ H turbine ⎠
3/ 4
(8)
Discussion It is difficult to achieve high efficiency in a pump-turbine during both the pump duty cycle and the turbine duty cycle. To achieve the highest possible efficiency, it is critical that both NSp and NSt are appropriate for the chosen design, e.g. radial flow centrifugal pump and radial-flow Francis turbine.
14-56 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-96 Solution
We are to apply conversion factors to prove a conversion factor.
Properties
We set g = 32.174 ft/s2 and assume water at density ρ = 1.94 slug/ft3.
Analysis
We convert NSt,US to NSt by dividing by g5/4 and ρ1/2, and then using conversion ratios to cancel all units, 1/ 2 (ω , rot/min )( bhp, hp ) 1 ⎛ 1 min ⎞ ⎛ slug ft ⎞ ⎛ 550 ft lbf ⎞ × ⎜ 60 s ⎟ ⎜ s 2 lbf ⎟ ⎜ s hp ⎟ 5/ 4 1/ 2 5/ 4 ⎝ ⎠⎝ ⎠ ⎝ ⎠ ( H , ft ) 1.94 slug/ft 3 ) ( 32.174 ft/s 2 ) ( 1/ 2
N St =
1/ 2
⎛ 2π rad ⎞ ⎜ s ⎟ ⎝ ⎠
N St ,US
N St = 0.02301N St ,US
Finally, the inverse of the above equation yields the desired conversion factor. Discussion As discussed in the text, some turbomachinery authors do not convert rotations to radians, introducing a confusing factor of 2π into the conversion.
14-97 Solution 14-108.
We are to calculate turbine specific speed for a given hydroturbine and compare to the guidelines of Fig.
Assumptions
1 The fluid is water at 20oC.
Properties
The density of water at 20oC is ρ = 998.0 kg/m3.
Analysis
First we convert the turbine rotation rate from rpm to radians/s,
ω = 180
Rotational speed:
rot ⎛ 2π rad ⎞ ⎛ 1 min ⎞ = 18.85 rad/s min ⎜⎝ rot ⎟⎠ ⎜⎝ 60 s ⎟⎠
With all other parameters already given in SI units, we calculate turbine specific speed,
ω ( bhp )
1/ 2
N St =
( ρ ) ( gH ) 1/ 2
5/ 4
=
(18.85 rad/s ) (119 ×106
W
)
1/ 2
( 998 kg/m ) ⎡⎣(9.81 m/s ) (105 m )⎤⎦ 3
1/ 2
2
1/ 2
5/ 4
⎛ m 2 kg ⎞ ⎜ 3 ⎟ ⎝ s ⋅W ⎠
= 1.12
Comparing to the turbine specific speed guidelines of Fig. 14-108, we see that this turbine should be of the Francis type, which it is. Discussion In fact, the turbine specific speed of this turbine is close to that which yields the maximum possible efficiency for a Francis turbine.
14-57 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-98 Solution 14-108.
We are to calculate turbine specific speed for a given hydroturbine and compare to the guidelines of Fig.
Assumptions
1 The fluid is water at 20oC.
Properties
The density of water at 20oC is ρ = 998.0 kg/m3.
Analysis
First we convert the turbine rotation rate from rpm to radians/s,
ω = 100
Rotational speed:
rot ⎛ 2π rad ⎞ ⎛ 1 min ⎞ = 10.47 rad/s min ⎜⎝ rot ⎟⎠ ⎜⎝ 60 s ⎟⎠
With all other parameters already given in SI units, we calculate turbine specific speed,
ω ( bhp )
N St =
( ρ ) ( gH ) 1/ 2
5/ 4
(10.47 rad/s ) (194 ×106 W )
1/ 2
1/ 2
=
1/ 2
( 998 kg/m ) ⎡⎣( 9.81 m/s ) ( 54.9 m )⎤⎦ 3
1/ 2
2
5/ 4
⎛ m 2 kg ⎞ ⎜ 3 ⎟ ⎝ s ⋅W ⎠
= 1.78
Comparing to the turbine specific speed guidelines of Fig. 14-108, we see that this turbine should be of the Francis type, which it is. Discussion The turbine specific speed of this turbine is close to the crossover point between a Francis turbine and a Kaplan turbine.
14-99 Solution 14-108.
We are to calculate turbine specific speed for a given hydroturbine and compare to the guidelines of Fig.
Assumptions
1 The fluid is water at 20oC.
Properties
The density of water at 20oC is ρ = 998.0 kg/m3.
Analysis
First we convert the turbine rotation rate from rpm to radians/s,
ω = 100
Rotational speed:
rot ⎛ 2π rad ⎞ ⎛ 1 min ⎞ = 10.47 rad/s min ⎜⎝ rot ⎟⎠ ⎜⎝ 60 s ⎟⎠
With all other parameters already given in SI units, we calculate turbine specific speed,
ω ( bhp )
1/ 2
N St =
( ρ ) ( gH ) 1/ 2
5/ 4
=
(10.47 rad/s ) ( 5.37 ×106
W
)
1/ 2
( 998 kg/m ) ⎡⎣( 9.81 m/s ) ( 9.75 m )⎤⎦ 3
1/ 2
2
1/ 2
5/ 4
⎛ m 2 kg ⎞ ⎜ 3 ⎟ ⎝ s ⋅W ⎠
= 2.57
Comparing to the turbine specific speed guidelines of Fig. 14-108, we see that this turbine should be of the Kaplan type, which it is. Discussion In fact, the turbine specific speed of this turbine is close to that which yields the maximum possible efficiency for a Kaplan turbine.
14-58 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-100 Solution
We are to calculate the specific speed of a turbine, and compare it to the normal range.
Analysis
We first calculate the nondimensional form of NSt,
ω ( bhp )
(12.57 rad/s ) ( 4.61×108 W )
1/ 2
1/ 2
N St =
ρ 1/ 2 ( gH )
5/ 4
=
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) ( 78.6 m )⎤⎦ 1/ 2
3
2
1/ 2
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
5/ 4
= 2.10
From Fig. 14-108, this is on the high end for Francis turbines – the designers may wish to consider a Kaplan turbine instead. From the conversion given in the text, NSt,US = 2.10 × 43.46 = 91.4. Alternatively, we can use the original dimensional data to calculate NSt,US,
( n, rpm )( bhp, hp ) N St ,US = 5/ 4 ( H , ft )
(120 rpm ) ( 6.18 ×105 hp )
1/ 2
1/ 2
=
⎛ ft ⎛ ⎞⎞ ⎜ 78.6 m ⎜ ⎟⎟ ⎝ 0.3048 m ⎠ ⎠ ⎝
5/ 4
= 91.3
Note that the actual values of brake horsepower and net head will differ from the values calculated here, because we have neglected irreversible losses; thus, the values of NSt and NSt,US may change slightly. Discussion We calculated the dimensional pump specific speed two ways as a check of our algebra, and (fortunately) the results agree (within round-off error).
14-101 Solution
We are to calculate the specific speed of a turbine, and compare it to the normal range.
Analysis
We first calculate the nondimensional form of NSt,
ω ( bhp )
(10.47 rad/s ) ( 2.98 ×107
1/ 2
N St =
ρ 1/ 2 ( gH )
5/ 4
=
(
998.0 kg/m3
) ( 1/ 2
)
W
)
1/ 2
⎡ 9.81 m/s 2 ( 38.05 m ) ⎤ ⎣ ⎦
1/ 2
5/ 4
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 1.10
From Fig. 14-108, this is in the range of the typical Francis turbine. From the conversion given in the text, NSt,US = 1.10 × 43.46 = 47.9. Alternatively, we can use the original dimensional data to calculate NSt,US,
( n, rpm )( bhp, hp ) N St ,US = 5/ 4 ( H , ft )
1/ 2
(100 rpm ) ( 4.00 ×104 hp ) = 5/ 4 (124.8 ft )
1/ 2
= 47.9
Note that the actual values of brake horsepower and net head will differ from the values calculated here, because we have neglected irreversible losses; thus, the values of NSt and NSt,US may change slightly. Discussion agree.
We calculated the dimensional pump specific speed two ways as a check of our algebra, and the results
14-59 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-102E Solution
We are to calculate the specific speed of a turbine, and compare it to the normal range.
Analysis
We first calculate the specific speed in customary US units,
( n, rpm )( bhp, hp ) 5/ 4 ( H , ft )
(120 rpm )( 202, 700 hp ) 5/ 4 (170.8 ft )
1/ 2
N St ,US =
1/ 2
=
= 87.5
From Fig. 14-108, this is on the high end of typical Francis turbines, on the border between Francis and Kaplan turbines. In nondimensional terms, NSt,US = 87.5 / 43.46 = 2.01. Note that the actual values of brake horsepower and net head will differ from the values calculated here, because we have neglected irreversible losses; thus, the values of NSt and NSt,US may change slightly. Discussion The actual values of brake horsepower and net head will differ from the values calculated here, because we have neglected irreversible losses; thus, the values of NSt and NSt,US may change slightly.
14-103 Solution
We are to calculate the specific speed of a turbine, and compare it to the normal range.
Analysis
We first calculate the nondimensional form of NSt, 1/ 2
N St =
ρ 1/ 2 ( gH )
(18.85 rad/s ) ( 2.46 ×108 W )
1/ 2
ω ( bhp )
5/ 4
=
1/ 2
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) ( 73.8 m )⎤⎦ 3
1/ 2
2
5/ 4
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 2.49
From Fig. 14-108, this is much higher than the typical Francis turbine – the designers should consider a Kaplan turbine instead. From the conversion given in the text, NSt,US = 2.49 × 43.46 = 108. Alternatively, we can use the original dimensional data to calculate NSt,US,
( n, rpm )( bhp, hp ) N St ,US = 5/ 4 ( H , ft )
(180 rpm ) ( 3.29 ×105 hp )
1/ 2
1/ 2
=
⎛ ft ⎛ ⎞⎞ ⎜ 73.8 m ⎜ ⎟⎟ ⎝ 0.3048 m ⎠ ⎠ ⎝
5/ 4
= 108
Note that the actual values of brake horsepower and net head will differ from the values calculated here, because we have neglected irreversible losses; thus, the values of NSt and NSt,US may change slightly. Discussion agree.
We calculated the dimensional pump specific speed two ways as a check of our algebra, and the results
14-60 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery
14-104 Solution
We are to determine a turbine’s efficiency and what kind of turbine is being tested.
Properties
The density of water at T = 20oC is 998.0 kg/m3.
Analysis
The turbine’s efficiency is calculated first,
η turbine =
bhp
ρ gHV
=
( 998.0 kg/m )( 3
2 ⎛ 3600 s ⎞ ⎛ kg ⋅ m ⎞ = 64.9% ⎜ ⎜ ⎟ 3 ⎟ 9.81 m/s 2 (15.0 m ) 17.0 m 3 /hr ⎝ hr ⎠ ⎝ W ⋅ s ⎠
450 W
)
(
)
After converting 1500 rpm to 157.1 rad/s, we calculate the nondimensional form of turbine specific speed,
ω ( bhp )
1/ 2
N St =
ρ 1/ 2 ( gH )
5/ 4
⎛ kg ⋅ m 2 ⎞ (157.1 rad/s )( 450 W ) 5/ 4 ⎜ 3 ⎟ 1/ 2 ( 998.0 kg/m3 ) ⎡⎣( 9.81 m/s2 ) (15.0 m )⎤⎦ ⎝ W ⋅ s ⎠
1/ 2
1/ 2
=
= 0.206
which we convert to customary US units, NSt,US = 0.206 × 43.46 = 8.94. This is most likely an impulse turbine (e.g., a Pelton wheel turbine). Discussion We could instead have used the dimensional units to directly calculate the turbine specific speed in customary US units. The result would be the same.
14-105 Solution
We are to scale up the model turbine tests to the prototype turbine.
Assumptions 1 The prototype and model are geometrically similar. 2 The tests are conducted under conditions of dynamic similarity. 3 The water is at the same temperature for both the model and the prototype (20oC).
The density of water at T = 20oC is 998.0 kg/m3.
Properties
Analysis We use the turbine scaling laws, starting with the head coefficient, and letting the model be turbine A and the prototype be turbine B, CH ,A =
gH A
ωA 2 DA 2
=
gH B
ωB 2 DB 2
= CH ,B
→
DA DB
ωB = ωA
HB HA
→
ωB = (1500 rpm )
1 50 m = 548 rpm 5 15.0 m
We then use the capacity coefficient to calculate the volume flow rate of the prototype, CQ ,A
VA VB = = = CQ ,B ωA DA 3 ωB DB3
→
ω ⎛D ⎞ VB = VA B ⎜ B ⎟ ωA ⎝ DA ⎠
3
→
3
548 rpm ⎛ 5 ⎞ VB = 17.0 m3 /hr = 776 m 3 /hr 1500 rpm ⎜⎝ 1 ⎟⎠
(
)
Finally, we use the power coefficient to calculate the brake horsepower of the prototype, CP ,A
bhp bhp = 3 A 5 = 3 B 5 = CP ,B ωA DA ωB DB
3
→
⎛ω ⎞ ⎛ D ⎞ bhpB = bhpA ⎜ B ⎟ ⎜ B ⎟ ⎝ ωA ⎠ ⎝ DA ⎠
5
3
→
⎛ 548 rpm ⎞ ⎛ 5 ⎞ VB = ( 450 W ) ⎜ ⎟ ⎜ ⎟ = 68, 500 W ⎝ 1500 rpm ⎠ ⎝ 1 ⎠ 5
Discussion All results are given to 3 significant digits, but we kept several extra digits in the intermediate calculations to reduce round-off errors.
14-61 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 14 Turbomachinery 14-106 Solution We are to compare the model and prototype efficiency and turbine specific speed to prove that they operate at homologous points. Properties
The density of water at T = 20oC is 998.0 kg/m3.
Analysis The model turbine’s efficiency was calculated in Problem 14-104 as 64.9%. We calculate the prototype turbine’s efficiency as
η turbine =
bhp
ρ gHV
=
( 998.0 kg/m )( 3
2 ⎛ 3600 s ⎞ ⎛ kg ⋅ m ⎞ ⎜ ⎟ = 64.9% ⎜ ⎟ 3 9.81 m/s 2 ( 50 m ) 776 m3 /hr ⎝ hr ⎠ ⎝ W ⋅ s ⎠
68,500 W
)
(
)
The efficiency of the prototype is the same as that of the model. Similarly, the turbine specific speed of the model turbine was calculated previously as 0.206. After converting 548 rpm to 57.36 rad/s, we calculate the nondimensional form of turbine specific speed for the prototype turbine,
ω ( bhp )
1/ 2
N St =
ρ 1/ 2 ( gH )
5/ 4
⎛ kg ⋅ m 2 ⎞ ( 57.36 rad/s )( 68,500 W ) = ⎜ 5 / 4 3 ⎟ 1/ 2 ( 998.0 kg/m3 ) ⎡⎣( 9.81 m/s2 ) (50 m )⎤⎦ ⎝ W ⋅ s ⎠ 1/ 2
1/ 2
= 0.206
Which is also the same as the previous calculation. Comparing to the results of Problem 14-104, we see that both ηturbine and Nst match between the model and the prototype. Thus, the model and the prototype operate at homologous points. Discussion Other nondimensional turbine parameters, like head coefficient, capacity coefficient, and power coefficient also match between model and prototype. We could instead have used dimensional units to calculate the turbine specific speed in customary US units.
14-107 Solution efficiency.
We are to estimate the actual efficiency of the prototype, and explain why it is higher than the model
Analysis
We apply the Moody efficiency correction equation,
ηturbine, prototype
⎛ D ≈ 1 − (1 − ηturbine, model ) ⎜ model ⎜ Dprototype ⎝
1/ 5
⎞ ⎟ ⎟ ⎠
1/ 5
⎛1⎞ = 1 − (1 − 0.649 ) ⎜ ⎟ ⎝5⎠
= 74.6%
This represents an increase of 74.6 – 64.9 = 9.7%. However, as mentioned in the text, we expect only about 2/3 of this increase, or 2(9.7%)/3 = 6.5%. Thus, our best estimate of the actual efficiency of the prototype is
ηturbine, prototype ≈ 64.9 + 6.5 = 71.4% There are several reasons for this increase in efficiency: The prototype turbine often operates at high Reynolds numbers that are not achievable in the laboratory. We know from the Moody chart that friction factor decreases with Re, as does boundary layer thickness. Hence, the influence of viscous boundary layers is less significant as turbine size increases, since the boundary layers occupy a less significant percentage of the flow path through the runner. In addition, the relative roughness (ε/D) on the surfaces of the prototype runner blades may be significantly smaller than that on the model turbine unless the model surfaces are micropolished. Finally, large full scale turbines have smaller tip clearances relative to blade diameter; therefore tip losses and leakage are less significant. Discussion The increase in efficiency between the model and prototype is significant, and helps us to understand why very large hydroturbines can be extremely efficient devices.
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Chapter 14 Turbomachinery 14-108 Solution We are to design a new hydroturbine by scaling up an existing hydroturbine. Specifically we are to calculate the new turbine diameter, volume flow rate, and brake horsepower. Assumptions 1 The new turbine is geometrically similar to the existing turbine. 2 Reynolds number effects and roughness effects are negligible. 3 The new penstock is also geometrically similar to the existing penstock so that the flow entering the new turbine (velocity profile, turbulence intensity, etc.) is similar to that of the existing turbine. Properties
The density of water at 20oC is ρ = 998.0 kg/m3.
Analysis Since the new turbine (B) is dynamically similar to the existing turbine (A), we are concerned with only one particular homologous operating point of both turbines, namely the best efficiency point. We solve Eq. 14-38b for DB, DB = DA
H B ⎛ nA ⎞ 110 m ⎛ 150 rpm ⎞ ⎜ ⎟ = (1.50 m ) ⎜ ⎟ = 2.0729 ≅ 2.07 m H A ⎝ nB ⎠ 90.0 m ⎝ 120 rpm ⎠
We then solve Eq. 14-38a for VB , 3
⎛ n ⎞ ⎛ D ⎞ ⎛ 120 rpm ⎞ ⎛ 2.0729 m ⎞ m3 m3 = 342.027 ≅ 342 VB = VA ⎜ B ⎟ ⎜ B ⎟ = 162 m3 /s ⎜ ⎟⎜ ⎟ s s ⎝ 150 rpm ⎠ ⎝ 1.50 m ⎠ ⎝ nA ⎠ ⎝ DA ⎠
(
3
)
Finally, we solve Eq. 14-38c for bhpB, 3
5
3
⎛ ρ ⎞ ⎛ n ⎞ ⎛ D ⎞ ⎛ 998.0 kg/m 3 ⎞ ⎛ 120 rpm ⎞ ⎛ 2.0729 m ⎞ bhpB = bhpA ⎜ B ⎟ ⎜ B ⎟ ⎜ B ⎟ = 132 MW ⎜ ⎟ ⎜ ⎟ = 340.62 MW ≅ 341 MW 3 ⎟⎜ ⎝ 998.0 kg/m ⎠ ⎝ 150 rpm ⎠ ⎝ 1.50 m ⎠ ⎝ ρ A ⎠ ⎝ nA ⎠ ⎝ DA ⎠
Discussion
5
To avoid round-off errors, we save several more significant digits for DB than are given in the final answer.
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Chapter 14 Turbomachinery
14-109 Solution
We are to compare the efficiency of two similar turbines, and discuss the Moody efficiency correction.
Analysis
We calculate the turbine efficiency for both turbines,
η turbine A =
η turbine B =
bhpA
ρ A gH AVA bhpB
ρ B gH BVB
=
=
( 998.0 kg/m )( 3
( 998.0 kg/m )( 3
⎛ kg ⋅ m 2 ⎞ = 92.5% ⎜ 3 ⎟ 9.81 m/s 2 ( 90 m ) 162 m 3 /s ⎝ W ⋅ s ⎠ 132 × 106 W
)
(
)
⎛ kg ⋅ m 2 ⎞ = 92.5% ⎜ 3 ⎟ 9.81 m/s 2 (110 m ) 342.027 m3 /s ⎝ W ⋅ s ⎠ 340.62 × 106 W
)
(
)
As expected, the two efficiencies are the same, since we have assumed dynamic similarity. However, total dynamic similarity may not actually be achieved between the two turbines because of scale effects (larger turbines generally have higher efficiency). The diameter of the new turbine is about 38% greater than that of the existing turbine, so the increase in efficiency due to turbine size should not be very significant. We verify this by using the Moody efficiency correction equation, considering turbine A as the “model” and B as the “prototype”, 1/ 5
ηturbine, B
Efficiency correction:
⎛D ⎞ ≈ 1 − (1 − ηturbine, A ) ⎜ A ⎟ ⎝ DB ⎠
1/ 5
⎛ 1.50 m ⎞ = 1 − (1 − 0.925 ) ⎜ ⎟ ⎝ 2.07 m ⎠
= 0.930
or 93.0%. So, the first-order correction yields a predicted efficiency for the larger turbine that is about half of a percent greater than the smaller turbine. However, as mentioned in the text, we expect only about 2/3 of this increase, or 2(0.5%)/3 = 0.3%. Thus, our best estimate of the actual efficiency of the prototype is ηturbine B ≈ 92.5 + 0.3 = 92.8% . Thus, our best estimate indicates that the new larger turbine will be slightly more efficient than its smaller brother, but the increase is only about 0.3%. Discussion If the flow entering the new turbine from the penstock were not similar to that of the existing turbine (e.g., velocity profile and turbulence intensity), we could not expect exact dynamic similarity.
14-110 Solution The turbine specific speed of two dynamically similar turbines is to be compared, and the most likely type of turbine is to be determined. Properties
The density of water at T = 20oC is ρ = 998.0 kg/m3.
Analysis
We calculate the dimensionless turbine specific speed for turbines A and B,
ωA ( bhpA )
1/ 2
N St ,A =
( ρA ) ( gH A ) 1/ 2
5/ 4
ωB ( bhpB )
=
( ρB ) ( gH B ) 1/ 2
5/ 4
W
)
1/ 2
1/ 2
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) (90.0 m )⎤⎦ 3
1/ 2
2
5/ 4
(12.57 rad/s ) ( 341×106 W )
1/ 2
1/ 2
N St ,B =
(15.71 rad/s ) (132 × 106
=
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) (110 m )⎤⎦ 3
1/ 2
2
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 1.19
1/ 2
5/ 4
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 1.19
We see that the turbine specific speed of the two turbines is the same. In customary US units, N St ,US,A = N St ,US,B = 43.46 N St = ( 43.46 )(1.19 ) = 51.6
From Fig. 14-108, both of these turbines are most likely Francis turbines. Discussion Since turbine A and turbine B operate at homologous points, it is no surprise that their turbine specific speeds are the same. In fact, if they weren’t the same, it would be a sure sign of an algebraic or calculation error.
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Chapter 14 Turbomachinery
Review Problems
14-111C Solution (a) True: As the gears turn, they direct a closed volume of fluid from the inlet to the outlet of the gear pump. (b) True or False: Rotary pumps can be either positive displacement or dynamic (an unfortunate use of terminology). As a positive-displacement pump, the rotors direct a closed volume of fluid from the inlet to the outlet of the rotary pump. As a dynamic pump, “rotary pump” is sometimes used in place of the more correct term, “rotodynamic pump”. (c) True: At a given rotational speed, the volume flow rate of a positive-displacement pump is fairly constant regardless of load because of the fixed closed volume. (d) False: Actually, the net head increases with fluid viscosity, because high viscosity fluids cannot penetrate the gaps as easily.
14-112C Solution
We are to discuss a water meter from the point of view of a piping system.
Analysis Although a water meter is a type of turbine, when analyzing pipe flow systems, we would think of the water meter as a type of minor loss in the system, just as a valve, elbow, etc. would be a minor loss, since there is a pressure drop associated with flow through the water meter. Discussion
In fact, manufacturers of water meters provide minor loss coefficients for their products.
14-113C Solution
We are to discuss the purpose of pump and turbine specific speeds.
Analysis Pump specific speed is used to characterize the operation of a pump at its optimum conditions (best efficiency point), and is useful for preliminary pump selection. Likewise, turbine specific speed is used to characterize the operation of a turbine at its optimum conditions (best efficiency point), and is useful for preliminary turbine selection. Discussion Pump specific speed and turbine specific speed are parameters that can be calculated quickly. Based on the value obtained, one can quickly select the type of pump or turbine that is most appropriate for the given application.
14-114C Solution
We are to discuss the definition and usefulness of a pump-turbine.
Analysis A pump-turbine is a turbomachine that can run both as a pump and as a turbine (by running in the opposite direction). A pump-turbine is used by some power plants for energy storage; specifically, water is pumped by the pump-turbine during periods of low demand for power, and electricity is generated by the pump-turbine during periods of high demand for power. Discussion We note that energy is “lost” both ways – when the pump-turbine is acting as a pump, and when it is acting as a turbine. Nevertheless, the energy storage scheme may still be cost-effective and profitable, in spite of the energy losses, because it may enable a power company to delay construction of costly new power-production facilities.
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Chapter 14 Turbomachinery
14-115 Solution turbines.
We are to prove a relationship between two dynamically similar pumps, and discuss its application to
Assumptions 1 The two pumps are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two pumps are not grossly different from each other in size, capacity, etc.). Analysis pumps,
Since the two pumps are dynamically similar, dimensionless pump parameter CH must be the same for both
CH ,A =
gH A
ω A 2 DA 2
= CH,B =
gH B
ω B 2 DB 2
→
ωA = ωB
H A DB H B DA
Similarly, dimensionless pump parameter CQ must be the same for both pumps, CQ ,A
VA VB = = C Q ,B ωA DA 3 ωB DB3
VB V
3
→
⎛ DB ⎞ ωA ⎜ ⎟ = ωB ⎝ DA ⎠
A
Combining the above two equations yields 3
⎛ DB ⎞ ⎜ ⎟ = ⎝ DA ⎠
H A DB VB H B DA VA
which reduces to 2
⎛ DB ⎞ ⎜ ⎟ = ⎝ DA ⎠
HA HB
VB V
A
1/ 4
→
⎛H ⎞ DB = DA ⎜ A ⎟ ⎝ HB ⎠
⎛ V ⎜⎜ B ⎝ VA
1/ 2
⎞ ⎟⎟ ⎠
Thus, we have eliminated the angular velocity as a parameter, and the relationship is proven. Since turbines follow the same affinity laws as pumps, the relationship also applies to two dynamically similar turbines. Discussion In like manner, we can eliminate other parameters through algebraic manipulations to scale a pump or turbine up or down.
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Chapter 14 Turbomachinery
14-116 Solution pumps.
We are to prove a relationship between two dynamically similar turbines, and discuss its application to
Assumptions 1 The two turbines are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two turbines are not grossly different from each other in size, capacity, etc.). Analysis Since the two turbines are dynamically similar, dimensionless turbine parameter CH must be the same for both turbines, CH ,A =
gH A
ω A 2 DA 2
= CH,B =
gH B
ω B 2 DB 2
ωA = ωB
→
H A DB H B DA
Similarly, dimensionless turbine parameter CP must be the same for both turbines, CP ,A
bhpA bhpB = = CQ ,B = 3 5 ρ Aω A DA ρ Bω B3 DB5
5
3
⎛ DB ⎞ ⎛ ω A ⎞ ρ A bhpB ⎜ ⎟ =⎜ ⎟ ⎝ DA ⎠ ⎝ ω B ⎠ ρ B bhpA
→
Combining the above two equations yields 5
⎛ DB ⎞ ⎛ H A ⎞ ⎜ ⎟ =⎜ ⎟ ⎝ DA ⎠ ⎝ H B ⎠
3/ 2
3
⎛ DB ⎞ ρ A bhpB ⎜ ⎟ ⎝ DA ⎠ ρ B bhpA
which reduces to 2
⎛ DB ⎞ ⎛ H A ⎞ ⎜ ⎟ =⎜ ⎟ ⎝ DA ⎠ ⎝ H B ⎠
3/ 2
ρ A bhpB ρ B bhpA
→
⎛H ⎞ DB = DA ⎜ A ⎟ ⎝ HB ⎠
3/ 4
1/ 2
⎛ ρA ⎞ ⎜ ⎟ ⎝ ρB ⎠
1/ 2
⎛ bhpB ⎞ ⎜ ⎟ ⎝ bhpA ⎠
Thus, we have eliminated the angular velocity as a parameter, and the relationship is proven. Since pumps follow the same affinity laws as turbines, the relationship also applies to two dynamically similar pumps. Discussion In like manner, we can eliminate other parameters through algebraic manipulations to scale a pump or turbine up or down.
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Chapter 14 Turbomachinery
Design and Essay Problems
14-117 [Also solved using EES on enclosed DVD] Solution We are to generate a computer application that uses the affinity laws to design a new pump that is dynamically similar to an existing pump. Assumptions 1 The two pumps are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two pumps are not grossly different from each other in size, capacity, etc.). Analysis bhpA =
First, we calculate the brake horsepower for pump A, 3 3 2 ρ AVA gH A ( 998.0 kg/m )( 0.00040 m /s )( 9.81 m/s ) (1.20 m ) ⎛ W ⋅ s3 = ⎜ 2 0.81 η pump,A ⎝ kg ⋅ m
⎞ ⎟ = 5.80 W ⎠
We use the affinity laws to calculate the new pump parameters. Using the given data, we first calculate the new pump diameter, 1/ 4
⎛H ⎞ DB = DA ⎜ A ⎟ ⎝ HB ⎠
⎛ VB ⎜⎜ ⎝ VA
1/ 2
⎞ ⎟⎟ ⎠
1/ 4
⎛ 120 cm ⎞ = ( 5.0 cm ) ⎜ ⎟ ⎝ 450 cm ⎠
1/ 2
⎛ 2400 cm3 /s ⎞ ⎜ ⎟ 3 ⎝ 400 cm /s ⎠
= 8.80 cm
Knowing DB, we calculate the rotation rate of the new pump, nB = nA
DA DB
1/ 2
⎛ HB ⎞ ⎜ ⎟ ⎝ HA ⎠
= (1725 rpm )
1/ 2
5.0 cm ⎛ 450 cm ⎞ 8.80 cm ⎜⎝ 120 cm ⎟⎠
= 1898 rpm
Knowing DB and ωB, we now calculate the required shaft power,
ρ bhpB = bhpA B ρA
5
3
3
⎛ DB ⎞ ⎛ ωB ⎞ 1226 kg/m3 ⎛ 8.80 cm ⎞ ⎛ 1898 rpm ⎞ ⎜ ⎟ ⎜ ⎟ = ( 5.80 W ) ⎜ ⎟ = 160 W 998.0 kg/m3 ⎜⎝ 5.0 cm ⎟⎠ ⎝ 1725 rpm ⎠ ⎝ DA ⎠ ⎝ ωA ⎠ 5
Our answers agree with those given. Now we are confident that we can apply our computer program to other design problems of similar nature, as in the next problem. Discussion To avoid round-off errors in the calculations, we saved several more significant digits for DB than are reported here.
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Chapter 14 Turbomachinery
14-118 We are to use a computer code and the affinity laws to design a new pump that is dynamically similar to an Solution existing pump. Assumptions 1 The two pumps are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two pumps are not grossly different from each other in size, capacity, etc.). 3 Both pumps operate at their BEP. Analysis Using our computer code, DB = 12.8 cm, nB = 1924 rpm, and bhpB = 235 W. We calculate pump specific speed for the new pump,
Pump B:
N Sp =
ωV 1/ 2
( gH )
3/ 4
⎛ 2π rad/s ⎞ 3 ⎟ 0.00367 m /s ⎝ 60 rpm ⎠
(
(1924 rpm ) ⎜ =
((
)
9.81 m/s 2 ( 5.70 m )
)
)
1/ 2
3/ 4
= 0.597
and we repeat the calculation for the existing pump,
Pump A:
N Sp =
ωV 1/ 2
( gH )
3/ 4
⎛ 2π rad/s ⎞ 3 ⎟ 0.00135 m /s ⎝ 60 rpm ⎠
(
(1500 rpm ) ⎜ =
( (9.81 m/s ) ( 2.10 m )) 2
3/ 4
)
1/ 2
= 0.597
Our answers agree, as they must, since the two pumps are operating at homologous points. From Fig. 14-73, we can see that these are most likely centrifugal pumps. Discussion code.
The pump performance parameters of pump B can be calculated manually to further verify our computer
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Chapter 14 Turbomachinery
14-119 Solution We are to generate a computer application that uses the affinity laws to design a new turbine that is dynamically similar to an existing turbine. Assumptions 1 The two turbines are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two turbines are not grossly different from each other in size, capacity, etc.). Analysis We use the affinity laws to calculate the new turbine parameters. Using the given data, we first calculate the new turbine diameter, DB = DA
H B ⎛ nA ⎞ 95.0 m ⎛ 150 rpm ⎞ ⎜ ⎟ = (1.40 m ) ⎜ ⎟ = 1.91 m H A ⎝ nB ⎠ 80.0 m ⎝ 120 rpm ⎠
We then solve Eq. 14-38a for VB , 3
⎛ n ⎞ ⎛ D ⎞ ⎛ 120 rpm ⎞ ⎛ 1.91 m ⎞ 3 VB = VA ⎜ B ⎟ ⎜ B ⎟ = 162 m3 /s ⎜ ⎟⎜ ⎟ = 328 m /s n D 150 rpm 1.40 m ⎝ ⎠ ⎝ ⎠ ⎝ A ⎠⎝ A ⎠
(
3
)
Finally, we solve Eq. 14-38c for bhpB, 3
5
3
⎛ ρ ⎞ ⎛ n ⎞ ⎛ D ⎞ ⎛ 998.0 kg/m3 ⎞ ⎛ 120 rpm ⎞ ⎛ 1.91 m ⎞ bhpB = bhpA ⎜ B ⎟ ⎜ B ⎟ ⎜ B ⎟ = 118 MW ⎜ ⎟ ⎜ ⎟ = 283 MW 3 ⎟⎜ ⎝ 998.0 kg/m ⎠ ⎝ 150 rpm ⎠ ⎝ 1.40 m ⎠ ⎝ ρ A ⎠ ⎝ nA ⎠ ⎝ DA ⎠ 5
Our answers agree with those given. Now we are confident that we can apply our computer program to other design problems of similar nature. Discussion To avoid round-off errors in the calculations, we saved several more significant digits for DB than are reported here.
14-120 Solution We are to use a computer program and the affinity laws to design a new turbine that is dynamically similar to an existing turbine. Assumptions 1 The two pumps are geometrically similar. 2 Reynolds number and roughness effects are not critical in the analysis (the two pumps are not grossly different from each other in size, capacity, etc.). Analysis Using our computer code, DB = 2.04 m, VB = 815 m3/s, and bhpB = 577 MW. We calculate turbine specific speed for the new turbine,
ωA ( bhpA )
1/ 2
N St ,A =
( ρA ) ( gH A ) 1/ 2
5/ 4
=
( 25.13 rad/s ) (11.4 ×106
W
)
1/ 2
1/ 2
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) ( 22.0 m )⎤⎦ 3
1/ 2
2
5/ 4
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 3.25
and for turbine B,
ωB ( bhpB )
1/ 2
N St ,B =
( ρ B ) ( gH B ) 1/ 2
5/ 4
=
( 21.99 rad/s ) ( 577 × 106
W
)
1/ 2
( 998.0 kg/m ) ⎡⎣( 9.81 m/s ) ( 95.0 m )⎤⎦ 3
1/ 2
2
1/ 2
5/ 4
⎛ kg ⋅ m 2 ⎞ ⎜ 3 ⎟ ⎝ W ⋅s ⎠
= 3.25
Our answers agree, as they must, since the two turbines are operating at homologous points. From Fig. 14-108, we see that these are most likely Kaplan turbines. Discussion
The turbine parameters of turbine B can be calculated manually to further verify our computer code.
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Chapter 14 Turbomachinery
14-121 Solution correction.
We are to compare efficiencies for two geometrically similar turbines, and discuss the Moody efficiency
Analysis
We calculate the turbine efficiency for both turbines,
η turbine A =
bhpA
ρ A gH AVA
=
( 998.0 kg/m )( 3
⎛ kg ⋅ m 2 ⎞ = 76.2% ⎜ 3 ⎟ 9.81 m/s 2 ( 22.0 m ) 69.5 m3 /s ⎝ W ⋅ s ⎠ 11.4 × 106 W
)
(
)
and
η turbine B =
bhpB
ρ B gH BVB
=
( 998.0 kg/m )( 3
⎛ kg ⋅ m 2 ⎞ = 76.2% ⎜ 3 ⎟ 9.81 m/s 2 ( 95.0 m ) 814.6 m3 /s ⎝ W ⋅ s ⎠ 577 × 106 W
)
(
)
where we have included an extra digit in the intermediate values to avoid round-off error. As expected, the two efficiencies are the same, since we have assumed dynamic similarity. However, total dynamic similarity may not actually be achieved between the two turbines because of scale effects (larger turbines generally have higher efficiency). The diameter of the new turbine is more than twice that of the existing turbine, so the increase in efficiency due to turbine size may be significant. We account for the size increase by using the Moody efficiency correction equation, considering turbine A as the “model” and B as the “prototype”,
Efficiency correction:
1/ 5
⎛ DA ⎞ ⎟ ⎝ DB ⎠
ηturbine, B ≈ 1 − (1 − ηturbine, A ) ⎜
1/ 5
⎛ 0.86 m ⎞ = 1 − (1 − 0.762 ) ⎜ ⎟ ⎝ 2.04 m ⎠
= 0.800
or 80.0%. So, the first-order correction yields a predicted efficiency for the larger turbine that is about four percent greater than the smaller turbine. However, as mentioned in the text, we expect only about 2/3 of this increase, or 2(0.800 – 0.762)/3 = 0.025 or 2.5%. Thus, our best estimate of the actual efficiency of the prototype is
ηturbine B ≈ 76.2 + 2.5 = 78.7% The higher efficiency of the new larger turbine is significant because an increase in power production of 2.5% can lead to significant profits for the power company. Discussion If the flow entering the new turbine from the penstock were not similar to that of the existing turbine (e.g., velocity profile and turbulence intensity), we could not expect exact dynamic similarity.
KJ
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Chapter 15 Computational Fluid Dynamics
Solutions Manual for Fluid Mechanics: Fundamentals and Applications by Çengel & Cimbala
CHAPTER 15 INTRODUCTION TO COMPUTATIONAL FLUID DYNAMICS
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Chapter 15 Computational Fluid Dynamics
Fundamentals, Grid Generation, and Boundary Conditions
15-1C Solution
We are to list the unknowns and the equations for a given flow situation.
Analysis There are only three unknowns in this problem, u, v, and P (orP′). Thus, we require three equations: continuity, x momentum (or x component of Navier-Stokes), and y momentum (or y component of Navier-Stokes). These equations, when combined with the appropriate boundary conditions, are sufficient to solve the problem. Discussion
The actual equations to be solved by the computer are discretized versions of the differential equations.
15-2C Solution
We are to define several terms or phrases and provide examples.
Analysis (a) A computational domain is a region in space (either 2-D or 3-D) in which the numerical equations of fluid flow are solved by CFD. The computational domain is bounded by edges (2-D) or faces (3-D) on which boundary conditions are applied. (b) A mesh is generated by dividing the computational domain into tiny cells. The numerical equations are then solved in each cell of the mesh. A mesh is also called a grid. (c) A transport equation is a differential equation representing how some property is transported through a flow field. The transport equations of fluid mechanics are conservation equations. For example, the continuity equation is a differential equation representing the transport of mass, and also conservation of mass. The Navier-Stokes equation is a differential equation representing the transport of linear momentum, and also conservation of linear momentum. (d) Equations are said to be coupled when at least one of the variables (unknowns) appears in more than one equation. In other words, the equations cannot be solved alone, but must be solved simultaneously with each other. This is the case with fluid mechanics since each component of velocity, for example, appears in the continuity equation and in all three components of the Navier-Stokes equation. Discussion
Students’ definitions should be in their own words.
15-3C Solution We are to discuss the difference between nodes and intervals and analyze a given computational domain in terms of nodes and intervals. Analysis Nodes are points along an edge of a computational domain that represent the vertices of cells. In other words, they are the points where corners of the cells meet. Intervals, on the other hand, are short line segments between nodes. Intervals represent the small edges of cells themselves. In Fig. P15-3 there are 6 nodes and 5 intervals on the top and bottom edges. There are 5 nodes and 4 intervals on the left and right edges. Discussion
We can extend the node and interval concept to three dimensions.
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Chapter 15 Computational Fluid Dynamics 15-4C Solution For a given computational domain with specified nodes and intervals we are to compare a structured grid and an unstructured grid and discuss. Analysis
We construct the two grids in the figure: (a) structured, and (b) unstructured.
(a)
(b)
There are 5 × 4 = 20 cells in the structured grid, and there are 36 cells in the unstructured grid. Discussion Depending on how individual students construct their unstructured grid, the shape, size, and number of cells may differ considerably.
15-5C Solution
We are to summarize the eight steps involved in a typical CFD analysis.
Analysis
We list the steps in the order presented in this chapter:
1. 2. 3. 4. 5. 6. 7. 8.
Specify a computational domain and generate a grid. Specify boundary conditions on all edges or faces. Specify the type of fluid and its properties. Specify numerical parameters and solution algorithms. Apply initial conditions as a starting point for the iteration. Iterate towards a solution. After convergence, analyze the results (post processing). Calculate global and integral properties as needed.
Discussion
The order of some of the steps is interchangeable, particularly Steps 2 through 5.
15-6C Solution
We are to explain why the cylinder should not be centered horizontally in the computational domain.
Analysis Flow separates over bluff bodies, generating a wake with reverse flow and eddies downstream of the body. There are no such problems upstream. Hence it is always wise to extend the downstream portion of the domain as far as necessary to avoid reverse flow problems at the outlet boundary. Discussion The same problems arise at the outlet of ducts and pipes – sometimes we need to extend the duct to avoid reverse flow at the outlet boundary.
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Chapter 15 Computational Fluid Dynamics 15-7C Solution
We are to discuss the significance of several items with respect to iteration.
Analysis (a) In a CFD solution, we typically iterate towards a solution. In order to get started, we make some initial conditions for all the variables (unknowns) in the problem. These initial conditions are wrong, of course, but they are necessary as a starting point. Then we begin the iteration process, eventually obtaining the solution. (b) A residual is a measure of how much our variables differ from the “exact” solution. We construct a residual by putting all the terms of a transport equation on one side, so that the terms all add to zero if the solution is correct. As we iterate, the terms will not add up to zero, and the remainder is called the residual. As the CFD solution iterates further, the residual should (hopefully) decrease. (c) Iteration is the numerical process of marching towards a final solution, beginning with initial conditions, and progressively correcting the solution. As the iteration proceeds, the variables converge to their final solution as the residuals decrease. (d) Once the CFD solution has converged, post processing is performed on the solution. Examples include plotting velocity and pressure fields, calculating global properties, generating other flow quantities like vorticity, etc. Post processing is performed after the CFD solution has been found, and does not change the results. Post processing is generally not as CPU intensive as the iterative process itself. Discussion
We have assumed steady flow in the above discussions.
15-8C Solution
We are to discuss how the iteration process is made faster.
Analysis (a) With multigridding, solutions of the equations of motion are obtained on a coarse grid first, followed by successively finer grids. This speeds up convergence because the gross features of the flow are quickly established on the coarse grid, and then the iteration process on the finer grid requires less time. (b) In some CFD codes, a steady flow is treated as though it were an unsteady flow. Then, an artificial time is used to march the solution in time. Since the solution is steady, however, the solution approaches the steady-state solution as “time” marches on. In some cases, this technique yields faster convergence. Discussion converge.
There are other “tricks” to speed up the iteration process, but CFD solutions often take a long time to
15-9C Solution We are to list the boundary conditions that are applicable to a given edge, and we are to explain why other boundary conditions are not applicable. Analysis We may apply the following boundary conditions: outflow, pressure inlet, pressure outlet, symmetry (to be discussed), velocity inlet, and wall. The curved edge cannot be an axis because an axis must be a straight line. The edge cannot be a fan or interior because such edges cannot be at the outer boundary of a computational domain. Finally, the edge cannot be periodic since there is no other edge along the boundary of the computational domain that is of identical shape (a periodic boundary must have a “partner”). The symmetry boundary condition merits further discussion. Numerically, gradients of flow variables in the direction normal to a symmetry boundary condition are set to zero, and there is no mathematical reason why the curved right edge of the present computational domain cannot be set as symmetry. However, you would be hard pressed to think of a physical situation in which a curved edge like that of Fig. P15-9 would be a valid symmetry boundary condition. Discussion Just because you can set a boundary condition and generate a CFD result does not guarantee that the result is physically meaningful.
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Chapter 15 Computational Fluid Dynamics 15-10C Solution
We are to discuss the standard method to test for adequate grid resolution.
Analysis The standard method to test for adequate grid resolution is to increase the resolution (by a factor of 2 in all directions if feasible) and repeat the simulation. If the results do not change appreciably, the original grid is deemed adequate. If, on the other hand, there are significant differences between the two solutions, the original grid is likely of inadequate resolution. In such a case, an even finer grid should be tried until the grid is adequately resolved. Discussion Keep in mind that if the boundary conditions are not specified properly, or if the chosen turbulence model is not appropriate for the flow being simulated by CFD, no amount of grid refinement is going to make the solution more physically correct.
15-11C Solution We are to discuss the difference between a pressure inlet boundary condition and a velocity inlet boundary condition, and we are to explain why both pressure and velocity cannot be specified on the same boundary. Analysis At a pressure inlet we specify the pressure but not the velocity. At a velocity inlet we specify the opposite – velocity but not pressure. To specify both pressure and velocity would lead to mathematical overspecification, since pressure and velocity are coupled in the equations of motion. When pressure is specified at a pressure inlet (or outlet), the CFD code automatically adjusts the velocity at that boundary. In a similar manner, when velocity is specified at a velocity inlet, the CFD code adjusts the pressure at that boundary. Discussion Since pressure and velocity are coupled, specification of both at a boundary would lead to inconsistencies in the equations of motion at that boundary.
15-12C Solution
We are to label all the boundary conditions to be applied to a computational domain.
Analysis The inlet is a velocity inlet. The outlet is a pressure outlet. All other edges that define the outer limits of the computational domain are walls. Finally, there are three edges that must be specified as interior. These are all labeled in the figure below.
Wall Pressure outlet
Wall Interior
V Wall
Velocity inlet Interior Wall
Wall
Interior Wall
Discussion correct.
It is critical that each boundary condition be specified carefully. Otherwise the CFD solution will not be
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Chapter 15 Computational Fluid Dynamics 15-13C Solution We are to analyze what will happen to inlet pressure and outlet velocity when a fan is turned on in the computational domain of the previous problem. Analysis Since the fan helps to push air through the channel, the inlet pressure will adjust itself so that less inlet pressure is required. In other words, the inlet pressure will decrease when the fan is turned on. Since the inlet velocity is the same in both cases, the mass flow rate (and volume flow rate since the flow is incompressible) must remain the same for either case. Therefore, outlet velocity will not change. Discussion It may seem at first glance that Vout should increase because of the fan, but in order to conserve mass, the outlet velocity cannot change. The solution is constrained by the specified inlet boundary condition. In a real physical experiment, there is no such restriction. The fan would cause the inlet pressure to decrease, the inlet velocity to increase, and the outlet velocity to increase.
15-14C Solution
We are to list and briefly describe six boundary conditions, and we are to give an example of each.
Analysis
In the chapter we list ten, so any six of these will suffice:
− − − − − − − − − −
Axis: Used in axisymmetric flows as the axis of rotation. Example: the axis of a torpedo-shaped body. Fan: An internal edge (2-D) or face (3-D) across which a sudden pressure rise is specified. Example: an axial flow fan in a duct. Interior: An internal edge (2-D) or face (3-D) across which nothing special happens – the interior boundary condition is used at the interface between two blocks. Example: all of the multiblock problems in this chapter, which require this boundary condition at the interface between any two blocks. Outflow: An outlet boundary condition in which the gradient of fluid properties is zero normal to the outflow boundary. Outflow is typically useful far away from the object or area of interest in a flow field. Example: the far field of flow over a body. Periodic: When the physical geometry has periodicity, the periodic boundary condition is used to specify that whatever passes through one face of the periodic pair must simultaneously enter the other face of the periodic pair. Example: in a heat exchanger where there are several rows of tubes. Pressure inlet: An inflow boundary in which pressure (but not velocity) is known and specified across the face. Example: the high pressure settling chamber of a blow-down wind tunnel facility. Pressure outlet: An outflow boundary in which pressure (but not velocity) is known and specified across the face. Example: the outlet of a pipe exposed to atmospheric pressure. Symmetry: A face over which the gradients of all flow variables are set to zero normal to the face – the result is a mirror image across the symmetry plane. Fluid cannot flow through a symmetry plane. Example: the mid-plane of flow over a circular cylinder in which the lower half is a mirror image of the upper half. Velocity inlet: An inflow boundary condition in which velocity (but not pressure) is known and specified across the face. Example: a uniform freestream inlet flow entering a computational domain from one side. Wall: A boundary through which fluid cannot pass and at which the no-slip condition (or a shear stress condition) is applied. Example: the surface of an airfoil that is being modeled by CFD.
Discussion this chapter.
There are additional boundary conditions used in CFD calculations, but these are the only ones discussed in
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Chapter 15 Computational Fluid Dynamics 15-15 Solution We are to sketch a structured and an unstructured grid near the airfoil surface, and discuss advantages and disadvantages of each. Analysis In either case it is wise to cluster cells close to the airfoil surface since we expect that a thin boundary layer will exist along the surface, and we need many tiny cells within the boundary layer to adequately resolve it. Some simple, coarse meshes are drawn in Fig. 1. We would certainly want much higher resolution for CFD calculations.
FIGURE 1 A coarse structured (a) and unstructured (b) grid. Notice that the cells are clustered (more fine) near the surface of the airfoil since there is likely to be large velocity gradients there (in the boundary layer).
(a)
(b)
The structured grid in Fig 1a is called a C-grid since it wraps around the airfoil like the letter “C”. The main advantage of the structured grid is that we can get high resolution near the surface with few cells. The main advantage of the unstructured grid is that it is somewhat easier to generate when the geometry is complicated (especially for highly curved surfaces). Furthermore, it is easier to transition between curved and straight edges with an unstructured grid. The main disadvantage of an unstructured grid is that more cells are required for the same spatial resolution. Discussion
There are numerous other ways to construct a grid around this airfoil.
15-16 Solution
We are to sketch a hybrid grid around an airfoil and explain its advantages.
Analysis We sketch a hybrid grid in the figure. Note that the grid is structured near the airfoil surface, but unstructured beyond the surface. The advantage of a hybrid grid is that it combines the advantages of both structured and unstructured grids. Near surfaces we can use a structured grid to finely resolve the boundary layer with a minimum number of cells, and away from surfaces we can use an unstructured grid so that we can rapidly expand the cell size. We can also more easily blend the grid into the edges of the computational domain with an unstructured grid. Discussion A structured grid is generally the best choice, but a hybrid grid is often a better option than a fully unstructured grid.
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Chapter 15 Computational Fluid Dynamics 15-17 Solution We are to sketch the blocking for a structured grid, sketch a coarse grid, and label all the boundary conditions to be applied to the computational domain. Analysis First of all, we recognize that because of symmetry, we can split the domain in half vertically. We construct four blocks around the half-cylinder to transform from round to square, and then we add simple rectangular blocks upstream and downstream of the cylinder (Fig. 1). There is a total of six blocks. Block 3
FIGURE 1 A possible blocking topology and coarse structured grid for a 2-D multiblock computational domain.
Block 1
Block 4
Block 2
Block 5
Block 6
Symmetry line
With the block structure of Fig. 1 no cells are highly skewed, and cells are clustered near the cylinder wall and the upper wall of the duct as desired. The bottom edge of the computational domain is a line of symmetry. The inlet is a velocity inlet. The outlet is a pressure outlet. The upper edge of the computational domain is a wall. The edges that define the cylinder are also walls. Finally, there are 5 edges that are specified as interior. These are all labeled in Fig. 2. Velocity inlet
Wall
Wall
Interior
FIGURE 2 Boundary conditions specified on each edge of a 2-D multiblock computational domain.
Interior
Symmetry
Wall
Pressure outlet
Symmetry
Discussion There are alternative ways to set up the blocking topology. For example, at the top we may define a thin block (Block 7) that stretches across the entire horizontal domain so that the boundary layer on the top wall of the channel can be more adequately resolved (Fig. 3). Block 3
FIGURE 3 An alternative blocking topology and coarse structured grid for the 2D multiblock computational domain. A seventh block is added at the top for better boundary layer resolution near the top plate.
Block 1
Block 2
Block 4
Block 5
Block 7
Block 6
Symmetry line
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Chapter 15 Computational Fluid Dynamics 15-18 Solution We are to sketch the blocking for a structured grid, sketch a coarse grid, and label all the boundary conditions to be applied to the computational domain. Analysis First of all, we recognize that because of symmetry, we can split the domain in half vertically. We construct four blocks inside the half-circle, and then we add blocks with one curved edge and three straight edges upstream and downstream of the cylinder (Fig. 1). Block 3 Block 2
FIGURE 1 The blocking and coarse structured grid for a 2-D multiblock computational domain.
Block 4
Block 1
Block 6
Block 5
Symmetry line
The setup of Fig. 1 contains six blocks. With this block structure, no cells are highly skewed, and cells are clustered near upper wall of the duct as desired. Cells are also clustered at the junctions between Blocks 1 and 2 and Blocks 4 and 6, where flow separation may occur. The bottom edge of the computational domain is a line of symmetry. The inlet is a velocity inlet. The outlet is a pressure outlet. The upper edge of the computational domain is a wall. The edges that define the cylinder are also walls. Finally, there are 5 edges that are specified as interior. These are all labeled in Fig. 2. Velocity inlet
FIGURE 2 Boundary conditions specified on each edge of a 2-D multiblock computational domain. Discussion
Wall Wall
Interior
Pressure outlet Interior
Interior Symmetry
Symmetry
There are of course, alternative ways to set up the blocking topology.
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Chapter 15 Computational Fluid Dynamics 15-19 Solution We are to modify an existing grid so that all blocks are elementary blocks. Then we are to verify that the total number of cells does not change. Analysis The right edge of Block 2 of Fig. 15-11b is split twice to accommodate Block 1. We therefore split Block 2 into three separate elementary blocks. Unfortunately, this process ends up splitting Block 6 twice, which in turn splits Block 4 twice. We end up with 12 elementary blocks as shown in Fig. 1. Block 11 (8×8)
Block 10 (8×4) Block 4 (5×8) Block 3 (5×4)
Block 5 (5×4)
Block 2 (5×8)
FIGURE 1 The blocking and coarse structured grid for a 2-D multiblock computational domain. Only elementary blocks are used in this grid.
Block 6 (5×8) Block 7 (5×4)
Block 1 (12×8) Block 9x (5×4)
Block 12 (8×4)
Block 8 (5×8)
We add up all the cells in these 12 blocks – we get a total of 464 cells. This agrees with the total of 464 cells for the original 6 blocks in the domain. Discussion Sometimes it is easier to create a grid with elementary blocks, even if the CFD code can accept blocks with split edges or faces.
15-20 Solution We are to modify an existing grid into a smaller number of non- elementary blocks, and we are to verify that the total number of cells does not change. Analysis We combine Blocks 2, 3, 4, and 5 of Fig. 15-10b. Together, these produce one structured grid that wraps around the square in the middle – there are still 5 i intervals, but now there are 48 j intervals. We end up with 3 nonelementary blocks, as shown in Fig. 1. Block 2 (5×48) Block 3 (8×16)
FIGURE 1 The blocking and coarse structured grid for a 2-D multiblock computational domain. Only elementary blocks are used in this grid.
Block 1 (12×8)
We add up all the cells in these 12 blocks – we get a total of 464 cells. This agrees with the total of 464 cells for the original 6 blocks in the domain. Discussion
Block 2 in Fig. 1 is called an O-grid (for obvious reasons).
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Chapter 15 Computational Fluid Dynamics 15-21 Solution We are to generate a computational domain and label all appropriate boundary conditions for one stage of a new heat exchanger design. Analysis We take advantage of the periodicity of the geometry. There are several ways to create a periodic grid for this flow. The simplest computational domain consists of a single flow passage between two neighboring tubes. We can make the periodic edge intercept anywhere on the front portion of the tube that we desire. We choose the lower surface for convenience and simplicity. The periodic computational domain is sketched in Fig. 1. Boundary conditions are also straightforward, and are labeled in Fig. 1. For a known inlet velocity we set the boundary condition at the left edge as a velocity inlet. The tube walls are obviously set as walls. The outlet can be set as either a pressure outlet or an outflow, depending on the provided information and how far the outlet region extends beyond the tubes. Finally, we set two pairs of translationally periodic boundaries, one fore and one aft of the tubes. We label them separately to avoid confusion. Velocity inlet
Wall Periodic “fore”
FIGURE 1 A periodic computational domain for a given geometry. Boundary conditions are also labeled.
Periodic “aft”
Wall
Pressure outlet or outflow
Wall
FIGURE 2 An alternative periodic computational domain for a given geometry. Boundary conditions are also labeled.
Velocity inlet Periodic “fore” Periodic “aft”
Pressure outlet or outflow
Wall
Discussion The fore and aft periodic edges are not horizontal in Fig. 1. This is not a problem since the periodic boundary condition is not restricted to horizontal or even to flat surfaces. An alternative, equally acceptable computational domain is shown in Fig. 2.
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Chapter 15 Computational Fluid Dynamics 15-22 Solution domain.
We are to sketch a structured multiblock grid with four-sided elementary blocks for a given computational
Analysis We choose the computational domain of Fig. 1 of the previous problem. Since all edges are straight, the blocking scheme can be rather simple. We sketch the blocking topology and apply a coarse mesh in Fig. 1 for the case in which the CFD code does not require the node distribution to be exactly the same on periodic pairs. FIGURE 1 The blocking topology and a coarse structured grid for a periodic computational domain. This blocking topology applies to CFD codes that allow a block’s edges to be split for application of boundary conditions, and do not require periodic edge pairs to have identical node distributions.
Block 1 Block 2
Block 3 Block 4
Block 5
Unfortunately, many CFD codes require that the node distribution on periodic pairs of edges be identical (the two edges of a periodic pair are “linked” in the grid generation process). In such a case, the grid of Fig. 1 would not be acceptable. Furthermore, although the edges of the blocks of Fig. 1 are not split with respect to adjacent blocks, the top edges of Block 1 and Block 3 are split with respect to the boundary conditions (part of the edge is periodic and part is a wall). Thus these blocks are not really elementary blocks after all. We construct a more elaborate blocking topology in Fig. 2 to correct these problems. The node distribution on the edges of each periodic pair are identical, at the expense of more complexity (7 instead of 5 blocks) and more cell skewness. Block 1 Block 2
Block 4
Block 3
FIGURE 2 The blocking topology and a coarse structured grid for a periodic computational domain. This blocking topology applies to CFD codes that require elementary blocks and require periodic edge pairs to have identical node distributions.
Block 5
Block 6 Block 7 Block 8
Discussion Some of the cells have moderate skewness with the blocking topology of Fig. 2, especially near the corners of Block 2 and Block 7 and throughout Block 5. A more complicated topology can be devised to reduce the amount of skewness.
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Chapter 15 Computational Fluid Dynamics 15-23 Solution We are to discuss why there is reverse flow in this CFD calculation, and then we are to explain what can be done to correct the problem. Analysis Reverse flow at an outlet is usually an indication that the computational domain is not large enough. In this case the rectangular heat exchanger tubes are inclined at 35o, and the flow will most certainly separate, leaving large recirculating eddies in the wakes. Anita should extend the computational domain in the horizontal direction downstream so that the eddies have a chance to “close” and the flow has a chance to re-develop into a flow without any reverse flow. Discussion In most commercial CFD codes a warning will pop up on the computer monitor whenever there is reverse flow at an outlet. This is usually an indication that the computational domain should be enlarged.
15-24 Solution We are to generate a computational domain and label all appropriate boundary conditions for two stages of a heat exchanger. Analysis We look for the smallest computational domain that takes advantage of the periodicity of the geometry. There are several ways to create a periodic grid for this flow. The simplest computational domain consists of a single flow passage between two neighboring tubes. We can make the periodic edge intercept anywhere on the fore and aft portions of the heat exchanger that we desire. We choose the periodic computational domain sketched in Fig. 1. Boundary conditions are also straightforward, and are labeled in Fig. 1. For a known inlet velocity we set the boundary condition at the left edge as a velocity inlet. The tube walls are obviously set as walls. The outlet can be set as either a pressure outlet or an outflow, depending on the provided information and how far the outlet region extends beyond the tubes. Finally, we set three pairs of translationally periodic boundaries, one fore, one mid, and one aft of the tubes. We label them separately to avoid confusion. Pressure outlet or outflow
Wall Velocity inlet
Wall Periodic “fore”
Periodic “mid”
Periodic “aft”
Wall
FIGURE 1 A periodic computational domain for a given geometry. Boundary conditions are also labeled. Discussion
Wall
Many other equally acceptable computational domains are possible.
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Chapter 15 Computational Fluid Dynamics 15-25 Solution domain.
We are to sketch a structured multiblock grid with four-sided elementary blocks for a given computational
Analysis We choose the computational domain of Fig. 1 of the previous problem. We sketch one possible elementary blocking topology in Fig. 1 for the case in which the CFD code requires the node distribution to be exactly the same on periodic pairs. We also assume that we cannot split one periodic edge and not its partner. The blocks are numbered. This topology has 16 elementary blocks.
FIGURE 1 The elementary blocking topology for a periodic computational domain. This blocking topology applies to CFD codes that do not allow a block’s edges to be split for application of boundary conditions, and requires periodic edge pairs to have identical node distributions.
1 5
2 3
4
6
16
7
12
8
15 14
11
9 10
13
Note that with this blocking topology we had to split the periodic “mid” boundary pair into two edges (the tops of blocks 6 and 7 and the bottoms of blocks 9 and 10). As long as both pairs of each segment are the same size and have the same number of nodes, this is not a problem. In the CFD code we would have to name each periodic pair separately, however. The block numbers are labeled. Notice that most of the blocks are nearly rectangular such that none of the computational cells would have to be highly skewed. Discussion This seems like a rather complicated blocking topology. It would require a bit of work to generate the grid. However, the time spent on developing a good grid is usually well worth the effort. By reducing the amount of cell skewness, we are able to speed up the CFD calculations and obtain more accurate results. This kind of topology also enables us to cluster cells near walls and wakes as needed.
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Chapter 15 Computational Fluid Dynamics FlowLab Problems
15-26 Solution We are to generate CFD solutions for external flow over a 2-D block. Specifically, we are to compare drag coefficient for various values of R/D (the extent of the outer boundary of the computational domain). In addition, we are to compare the calculated value of CD with experiment. TABLE 1 Drag coefficient as a function of the Assumptions 1 The flow is two-dimensional and incompressible. 2 The normalized extent of the flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the computational domain for turbulent mean. flow over a rectangular block. Properties kg/m⋅s.
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5
Analysis (a) The Reynolds number is Re =
3 ρVD (1.225 kg/m ) ( 2.0 m/s )( 0.10 m ) = = 1.37 ×104 μ 1.7894 × 10−5 kg/m ⋅ s
R/D 5 10 20 50 100 200 500
CD 1.81927 1.50662 1.41076 1.36723 1.35282 1.34671 1.34408
Experimental data indicate that the drag coefficient for this body is CD ≈ 1.9 at Reynolds numbers greater than 104 (see Chap. 11). (b) The CFD code is run for eight values of R, all else being equal. CD is tabulated as a function of R/D in Table 1, and plotted in Fig. 2. As the extent of the computational domain grows in size, the drag coefficient decreases steadily, but levels off to three significant digits of precision by R/D ≈ 200. Thus, a computational domain extent of R/D ≈ 100 is sufficient to achieve 1.9 independence of CD. We report a final value of CD = 1.34. (c) There are several possible reasons for the discrepancy between the 1.8 calculated value of CD (1.34) and the experimentally obtained value of CD (about 1.9). First of all, the actual flow is most likely unsteady, with 1.7 vortices being shed into the wake, whereas we are simulating a steady CD 1.6 flow. In addition, the unsteady shedding of vortices renders the flow no longer symmetric about the x axis, whereas we are forcing our flow to be symmetric. Furthermore, the grid resolution may not be adequate to 1.5 achieve grid independence (this is checked in the following problem). 1.4 Finally (and most importantly), we are using a turbulence model to simulate this flow field. The CFD solution we obtain is only as good as 1.3 the degree to which the turbulence model correctly models the physics of 0 1 the turbulence. As discussed in the text, no turbulence model is 10 10 102 103 1 10 100 1000 universally valid for all types of turbulent flows. Discrepancies between R/D experiment and CFD will always exist regardless of how fine the grid or FIGURE 1 how large the extent of the computational domain. Drag coefficient plotted as a function of the normalized extent of the computational (d) Streamlines near the body are plotted for R/D = 5 and 500 in Fig. 2. domain for turbulent flow over a rectangular We notice that the streamlines for the R/D = 5 case are more tight around block. the body compared to those for R/D = 500. This is most likely due to interference from the outer edges of the computational domain, which are too close for the R/D = 5 case.
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Chapter 15 Computational Fluid Dynamics
FIGURE 2 Streamlines for steady, incompressible, two-dimensional, turbulent flow over a rectangular block at R/D = (a) 5, (b) 500. Only the upper half of the flow is simulated
Block
Block (a) R/D = 5
(b) R/D = 500
Discussion Newer versions of FlowLab may give slightly different results. It is perhaps surprising how far away the edges of the computational domain must be in order for them not to influence the flow field around the body. No attempt was made here to optimize the grid, and although the comparison between various grid extents is valid, the solution itself may not be grid independent; the actual value of drag coefficient may not be correct due to lack of grid resolution. Grid independence is analyzed in the next problem.
15-27 Solution We are to test for grid independence by running various values of grid resolution, and we are to examine the effect of grid resolution on drag coefficient. Assumptions 1 The flow is two-dimensional and incompressible. 2 The flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. TABLE 1 Drag coefficient as a function of number of 3 Properties The fluid is air with ρ = 1.225 kg/m and μ = 1.7894 × cells in the computational domain for -5 10 kg/m⋅s. turbulent flow over a rectangular block. Analysis We run the CFD code for several grid resolutions, and Number of cells CD we list drag coefficient as a function of total number of computational 3120 1.45505 cells in Table 1. Although some discrepancies exist in the third or 10400 1.35249 fourth digit, the drag coefficient levels off to three significant digits of 12480 1.34893 precision by the sixth row of the table. Further grid refinement may be 18720 1.33839 necessary. We report a final value of CD = 1.33 (to three significant 23920 1.33595 digits of precision for comparison with experimental results). This 26000 1.33379 result differs from the experimentally obtained value of CD = 1.9 by about 30%. We have achieved grid independence to three digits, and thereby eliminate lack of grid resolution as a source of the discrepancy. The other reasons for the discrepancy between CFD and experiment remain, regardless of how fine the grid. Namely, nonuniversality of the turbulence model, unsteadiness, and nonsymmetry. Discussion Newer versions of FlowLab may give slightly different results. This exercise illustrates that grid refinement does not necessarily lead to improved CFD predictions.
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Chapter 15 Computational Fluid Dynamics 15-28 Solution We are to repeat the CFD calculation of drag coefficient around a rectangular block for two other fluids – water and kerosene, and we are to compare our results to those obtained using air as the fluid. Assumptions 1 The flow is two-dimensional and incompressible. 2 The flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties The density and viscosity of the default air are ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s. The density and viscosity of liquid water at T = 15oC are 998.2 kg/m3 and 1.003 × 10-3 kg/m⋅s. The density and viscosity of kerosene at T = 15oC are 780.0 kg/m3 and 2.40 × 10-3 kg/m⋅s.
TABLE 1 Drag coefficient as a function of fluid type for the case of turbulent flow over a rectangular block. In all cases, the Reynolds number is the same. Fluid Air Water Kerosene
Re 1.37 × 104 1.37 × 104 1.37 × 104
CD 1.34344 1.34343 1.34343
Analysis A comparison of the CFD calculations for all three fluids is given in Table 1. The drag coefficient is identical to about five digits of precision. We conclude that for incompressible flow without free surface effects, the Reynolds number is the critical parameter; the type of fluid is irrelevant provided that the Reynolds number is the same. This reinforces what we learned about dimensional analysis in Chap. 7. Discussion Newer versions of FlowLab may give slightly different results. Some incompressible CFD codes work with normalized variables from the start, requiring input of a Reynolds number instead of dimensional quantities such as velocity, density, and viscosity.
15-29 Solution discuss.
We are to generate CFD solutions for drag coefficient as a function of Reynolds number and compare and
Assumptions 1 The flow is two-dimensional and incompressible. 2 The flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties kg/m⋅s.
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5
TABLE 1 Drag coefficient as a function of Reynolds number for turbulent flow over a rectangular block. Re 10000 50000 100000 500000 1.00E+06 3.00E+06 5.00E+06
CD 1.30788 1.40848 1.42215 1.43065 1.43194 1.43296 1.43431
Analysis We compare six cases in Table 1. We see that CD levels off to a value of 1.43 to three digits of precision for Re greater than about 5 × 105. Thus, we have achieved Reynolds number independence, although the required Reynolds number is somewhat larger than that required experimentally. The last two cases are peculiar in that the Mach numbers are well beyond the incompressible limit (around 0.3) since the speed of sound in air at room temperature is around 340 m/s. However, even though these flows are unphysical, the CFD code is run as incompressible, and is not “aware” of this problem since the speed of sound is treated as infinite in an incompressible flow solver. The comparison with Re is still valid since we can use any incompressible fluid for the calculations, as illustrated in the previous problem. Discussion Newer versions of FlowLab may give slightly different results. Reynolds number independence checks are not always as simple as that shown here, because as Re increases, boundary layer thicknesses tend to decrease, requiring a finer mesh near walls. In the present problem this is not really an issue because the flow separates at the sharp edges of the block, and boundary layer thickness is not an important parameter in calculation of the drag.
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Chapter 15 Computational Fluid Dynamics 15-30 Solution We are to generate CFD solutions using several turbulence models, and we are to compare and discuss the results, particularly the drag coefficient. Assumptions 1 The flow is two-dimensional and incompressible. 2 The flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s.
Analysis All cases are run at the same Reynolds number, namely 1.37 × 104. We compare drag coefficient for all four cases in Table 1. We see that CD depends greatly on turbulence model. Most of the models underpredict the drag coefficient by about 30%, but the k-ω model overpredicts CD by more than 33%. In terms of percentage error, the Reynolds stress model is closest to the experimental value of 1.9. TABLE 1 Drag coefficient as a function of turbulence model for flow over a rectangular block. The error is in comparison to the experimental value of 1.9. Turbulence model k-ε (2 eq.) k-ω (2 eq.) Spallart-Allmaras (1 eq.) Reynolds stress model (5 eq.)
CD 1.34342 2.536939 1.360602 1.385374
Error (%) -29.5% 33.5% -28.4% -27.1%
Discussion Newer versions of FlowLab may give slightly different results. Turbulence models involve semi-empirical analysis, curve-fits, and simplifications, and no turbulence model is best for every kind of fluid flow. It is not always clear which turbulence model to use for a given problem.
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Chapter 15 Computational Fluid Dynamics 15-31 Solution We are to run CFD simulations of flow over a 2-D rectangular block with various values of L/D. We are to compare and discuss streamlines and drag coefficient, and we are to compare our CFD results to experiment. 2.6 2.4
Assumptions 1 The flow is two-dimensional and incompressible. 2 The flow is symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties 10-5 kg/m⋅s.
The fluid is air with ρ = 1.225 kg/m and μ = 1.7894 × 3
Experiment
2.2
CD
2 1.8 1.6
Analysis (a) Drag coefficient is tabulated as a function of L/D in Table 1. These data are also plotted in Fig. 1, along with experimental data from Chap. 11. We see that the calculations are consistently lower than the experimental values for the smaller lengths, but the agreement is very good at L/D = 3 (as high as the table goes). While the experimentally obtained drag coefficient peaks at L/D = 0.5, the CFD calculations predict that CD decays continually with L/D.
1.4
CFD
1.2 1 0
1
2 3 L/D
4
5
FIGURE 1 Drag coefficient as a function of L/D for turbulent flow over a rectangular block.
(b) For several values of L/D, we plot streamlines near the block (Fig. 2). We can see why the drag coefficient decreases as L/D increases. Namely, as the block length increases, the large recirculating flow regions in the wake (wake eddies) decrease in size (particularly in the vertical direction). In all cases, the flow separates at the sharp corner of the blunt face, but as L increases, the flow has TABLE 1 more time to flatten out along the upper and lower walls of the block, leading to Drag coefficient as a function of wake eddies of reduced thickness. Indeed, by L/D = 3.0 (Fig. 2f) the separated L/D for turbulent flow over a flow along the upper and lower walls appears to reattach just upstream of the rectangular block. back of the block. The wake eddies in this case are much thinner than those of the shorter blocks (compare Fig. 2b and 2e for example). Another way to express this L/D CD is to say that the longer blocks are more “streamlined” (in a gross sense of the 0.1 1.53997 word) than are the shorter blocks, and therefore have less drag. Experiments 0.5 1.50858 show that the drag is highest at L/D = 0.5. Apparently the vortex shedding 1 1.46466 process is strong for this case, leading to high drag. Our steady, symmetric CFD 1.5 1.42077 model is not able to simulate the unsteady features of the actual flow. 2 1.36825 3 1.26405 5 1.09386 (c) There are many possible reasons for the discrepancy between CFD calculations and experiment. We are modeling the problem as a steady flow that is symmetric about the axis, but experiments reveal that flow over bluff bodies like these oscillate and shed vortices – the flow is neither steady nor symmetric. Furthermore, we are using a turbulence model. As discussed previously, turbulence models are not universal, and may not be applicable to the present problem. A DNS or LES simulation would be required to correctly model the unsteady turbulent eddies.
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Chapter 15 Computational Fluid Dynamics
(a) L/D = 0.1
(b) L/D = 0.5
(c) L/D = 1.0
(d) L/D = 1.5
(e) L/D = 2.0
(f) L/D = 3.0
FIGURE 2 Streamlines for steady, incompressible, two-dimensional, turbulent flow over a rectangular block with various values of block length to block height: L/D = (a) 0.1, (b) 0.5, (c) 1.0, (d) 1.5, (e) 2.0, and (f) 3.0.
Discussion Newer versions of FlowLab may give slightly different results. CFD issues such as grid resolution and the extent of the computational domain do not contribute appreciably to the discrepancy here, since we have set up the computational domain based on results of similar previous problems. As L/D increases, the discrepancy between CFD results and experiment gets smaller. This can be explained by the fact that the shed vortices are reduced in strength as L/D increases – the steady, symmetric CFD simulation thus becomes more physically correct with increasing L/D.
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Chapter 15 Computational Fluid Dynamics 15-32 Solution We are to generate CFD solutions for external flow over a cylindrical block. Specifically, we are to compare drag coefficient for various values of R/D (the extent of the outer boundary of the computational domain). In addition, we are to compare the calculated value of CD with experiment. TABLE 1 Assumptions 1 The flow is incompressible. 2 The flow is Drag coefficient as a function of the axisymmetric about the x axis. 3 The flow is turbulent, but steady in normalized extent of the computational the mean. domain for turbulent flow over a cylindrical block. 3 Properties The fluid is air with ρ = 1.225 kg/m and μ = 1.7894 × R/D CD 10-5 kg/m⋅s. 5 1.0704 10 1.04053 Analysis 20 1.03567 (a) The Reynolds number based on cylinder diameter is the same as 50 1.02155 that of Problem 15-26, namely, 100 1.00781 3 200 0.998819 ρVD 1.225 kg/m ( 2.0 m/s )( 0.10 m ) 4 500 0.992311 Re = = = 1.37 ×10 1.7894 × 10−5 kg/m ⋅ s μ
(
)
Experimental data indicate that the drag coefficient for this body is CD ≈ 0.90 at Reynolds numbers greater than 104 (see Chap. 11). (b) The CFD code is run for eight values of R, all else being equal. CD is tabulated as a function of R/D in Table 1, and plotted in Fig. 2. As the extent of the computational domain grows in size, the drag coefficient decreases steadily, but is trying to level off by R/D ≈ 500. Thus, a computational domain extent of R/D ≈ 500 or more is needed to achieve independence of CD. We report a final value of CD = 0.99. Unfortunately, the program does not allow for R/D values greater than 500. (c) There are several possible reasons for the discrepancy between the calculated value of CD and the experimentally obtained value of CD (about 0.90). First of all, the actual flow is most likely unsteady, with vortices being shed into the wake, whereas we are simulating a steady flow. In addition, the unsteady shedding of vortices renders the flow no longer axisymmetric about the x axis, whereas we are forcing our flow to be axisymmetric. Furthermore, the grid resolution may not be adequate to achieve grid independence. Finally (and most importantly), we are using a turbulence model to simulate this flow field. The CFD solution we obtain is only as good as the degree to which the turbulence model correctly models the physics of the turbulence. As discussed in the text, no turbulence model is universally valid for all types of turbulent flows. Discrepancies between experiment and CFD will always exist regardless of how fine the grid or how large the extent of the computational domain.
1.08 1.07 1.06 1.05 1.04 CD 1.03 1.02 1.01 1 0.99 0.98 0
100 200 300 400 500 R/D
FIGURE 1 Drag coefficient plotted as a function of the normalized extent of the computational domain for turbulent flow over a rectangular block.
(d) Streamlines near the body are plotted for R/D = 5 and 500 in Fig. 2. There is surprisingly little observable difference between these two extreme cases. Compared to the 2-D case, it appears that the axisymmetric case requires a greater computational domain extent. This is surprising, because in an axisymmetric flow field, the fluid can flow around the body in all directions, not just over the top and bottom as in 2-D flow. As fluid moves away from the body, the radius also increases there, and more mass can flow through that radial location compared to the 2-D case. In other words an axisymmetric flow is less “confined” or “constrained” than a corresponding two-dimensional flow. Thus, we might have expected the opposite behavior. We compare our predicted drag coefficient (0.99) with that of experiment (0.90). The discrepancy is only about 10% − much better agreement than the two-dimensional case. The reasons for this improvement is not clear. Axisymmetric flows tend to be less unsteady, and the vortices they shed are generally less coherent and weaker. This may contribute to some of the improvement. It may be merely fortuitous that the turbulence model yields better results for the axisymmetric case as compared to the 2-D case.
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Chapter 15 Computational Fluid Dynamics
FIGURE 2 Streamlines for steady, incompressible, axisymmetric, turbulent flow over a rectangular block at R/D = (a) 5, (b) 500. Only one slice of the flow is shown.
Block
Block (a) R/D = 5
(b) R/D = 500
Discussion Newer versions of FlowLab may give slightly different results. The far field boundaries must be quite far away to achieve results that are independent of the extent of the boundary.
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Chapter 15 Computational Fluid Dynamics 15-33 Solution We are to generate CFD solutions for several grid resolutions to test for grid independence for flow through a diffuser. Specifically, we are to compare streamlines and pressure difference at each level of grid resolution. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s.
Analysis (a) Streamlines are plotted for six grid resolution cases in Fig. 1. (a)
FIGURE 1 Streamlines for steady, incompressible, axisymmetric, turbulent flow through a diffuser at various levels of grid resolution; the number of cells is (a) 880, (b) 3520, (c) 4950, (d) 8800, (e) 13750, and (f) 19800.
(b)
x (c)
x (d)
x (e)
x
At the very coarse grid resolution, the streamlines are not well defined (Fig. 1a), and the calculation is not reliable. As grid resolution improves, details of the flow separation region become more refined – the boundary layer is unable to remain attached in such a strong adverse pressure gradient. From these plots, it appears that grid independence has been achieved by about the fourth case (Fig. 1d), beyond which there is no noticeable change in the shape of the streamlines. We note that simulation of flow separation and separation bubbles is often a very difficult task for a CFD program. In this particular problem we must use an extremely fine grid in order to resolve the details of the flow separation. (b) ΔP is tabulated as a function of cell count in the table for the case with θ = 20o. As grid resolution improves, ΔP increases, and becomes independent of grid resolution by the fourth or fifth mesh. This is also seen in Fig. 2 where ΔP is plotted as a function of cell count.
x (f)
x -31.5 -32 -32.5
ΔP (Pa)
-33 -33.5 -34 -34.5
cell count 880 3520 4950 8800 13750 19800
ΔP (Pa) -34.5622 -32.7267 -32.4183 -32.2021 -32.1162 -32.1099
-35
102 100
103 104 105 1000 10000 100000 Number of cells
FIGURE 2 Pressure difference as a function of the number of cells in the computational domain. Turbulent flow through an axisymmetric diffuser.
Discussion The unphysical-looking streamlines at the very coarse grid resolution (Fig. 1a) are due to interpolation errors when the CFD code calculates contours of constant stream function. Notice that even though there is gross flow separation in this diffuser, there is still a pressure recovery through the diffuser (Pin is less than Pout). A better design (with even higher pressure recovery) would use a smaller diffuser angle so as to avoid flow separation along the diffuser wall. The outlet of the computational domain is reasonably far downstream (several pipe diameters) to avoid reverse flow at the outlet.
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Chapter 15 Computational Fluid Dynamics 15-34 Solution We are to compare the pressure drop and the pressure distribution at the outlet of a diffuser in a round pipe with two different outlet conditions: pressure outlet and outflow. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s.
Analysis For the present case (outflow boundary condition), the pressure difference is ΔP = Pin – Pout = −32.1096 Pa, which agrees to four digits of precision with the result of the previous problem at the same grid resolution, for which ΔP = Pin – Pout = −32.1090 Pa. Thus we conclude that the outlet boundary condition has negligible effect on this flow field. For the case with the pressure outlet boundary condition, the static pressure at the outlet of the computational domain is forced to be constant (zero gage pressure in the calculations of the previous problem). In the present case however, the outflow boundary condition does not fix static pressure – rather, it forces flow variables to level off as they approach the outlet boundary. We find that P varies by less than 0. percent across the boundary. Thus, even though P is not forced to be constant along the outflow boundary, it turns out to be nearly constant anyway. Discussion Newer versions of FlowLab may give slightly different results. With a velocity inlet and an outflow outlet, we do not fix the value of pressure at either boundary. Instead, the CFD code assigns P = 0 gage pressure at some (arbitrary) location in the flow field (the default location is the origin). Even though the inlet and outlet pressures differ significantly between the two cases, ΔP is identical to within four significant digits of precision. These results verify the statement we made in Chap. 9: For incompressible flow, it is not pressure itself which is important to the flow field, but rather pressure differences. We conclude that the differences between pressure outlet and outflow are small, provided that the outlet boundary is far enough away from the region of interest in the flow field.
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Chapter 15 Computational Fluid Dynamics 15-35 Solution We are to generate CFD solutions for flow through a diffuser at various values of diffuser half-angle θ. Specifically, we are to compare streamlines and pressure difference for each case, and we are to determine the maximum value of θ that achieves the stated design objectives. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s.
Analysis (a) Streamlines are plotted in Fig. 1 for all twelve cases.
FIGURE 1 Streamlines for steady, incompressible, axisymmetric, turbulent flow through a diffuser at various values of diffuser halfangle: θ = (a) 5o, (b) 7.5o, (c) 10o, (d) 12.5o, (e) 15o, (f) 17.5o, (g) 20o, (h) 25o, (i) 30o, (j) 45o, (k) 60o, and (l) 90o. (The latter case is a sudden expansion.)
(a) θ = 5o
x
(b) θ = 7.5o
x
(c) θ = 10o
x
(d) θ = 12.5o
x
(e) θ = 15o
x
(f) θ = 17.5o
x
(g) θ = 20o
x
(h) θ = 25o
x
(i) θ = 30o
x
(j) θ = 45o
x
(k) θ = 60o
x
(l) θ = 90o
x
At θ = 5o and 7.5o, the flow does not separate along the diffuser wall (Figs. 1a and 1b), although separation appears imminent near the downstream corner of the diffuser for the latter case. To avoid flow separation, Barb should recommend a diffuser half-angle of 7.5o or less. As θ increases, the boundary layer is unable to remain attached in the adverse pressure gradient, and the flow separates. A very small separation bubble is apparent at θ = 10o (Fig. 1c). As θ continues to increase, the separation bubble grows in size, and the separation point moves upstream, closer and closer to the upstream corner of the diffuser (compare Figs. 1d through 1g). By θ = 25o (Fig. 1h), the flow separates very close to the start of the diffuser. From this point on, the diffuser angle is so sharp that the flow separates right at the upstream corner of the diffuser. The streamline patterns reveal that the separation bubble continues to grow in size as θ increases (Figs. 1i through 1j). Beyond θ ≈ 45o however, the streamline pattern changes very little in the separation bubble (compare Figs. 1k and 1l).
-10.0
-20.0
ΔP (Pa) -30.0
-40.0
-50.0 0
50
θ (degrees)
100
FIGURE 2 Pressure difference from inlet to outlet as a function of diffuser half-angle θ.
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Chapter 15 Computational Fluid Dynamics
(b) ΔP is tabulated as a function of diffuser half-angle in the table (four extra cases are solved for improved clarity). As θ increases, Pin increases, reflecting the effect of the larger separation bubble. Physically, we achieve less and less pressure recovery as the separation bubble grows. We see that ΔP flattens out at high values of θ, becoming nearly independent of θ beyond θ ≈ 60o. This is also seen in Fig. 2 where Pin is plotted as a function of diffuser half-angle. There is a sharp rise in ΔP beyond 30o. The reason for this is not certain, but is probably related to the fact that when θ is greater than about 30o, the flow separates right at the upstream corner. The pressure rise is greater than 40 Pa for all angles below 15o. Thus, to ensure a pressure recovery of at least 40 pascals, Barb should recommend a diffuser half-angle of 12.5o or less. Discussion The results here are for an older version of FlowLab, and the results at 20o are therefore not exactly the same as those of the previous problem. Newer versions of FlowLab would give slightly different results, and the agreement would be exact. The outlet of the computational domain is reasonably far downstream (several pipe diameters) to avoid reverse flow at the outlet. Notice that even for the case of a sudden expansion (θ = 90o) there is still a pressure recovery through the diffuser (Pin is less than Pout).
θ 5 7.5 10 12.5 15 17.5 20 25 30 32.5 35 37.5 45 60 75 90
ΔP -49.1371 -47.7787 -44.9927 -42.4013 -39.6981 -37.6431 -36.0981 -32.7173 -29.9919 -23.2118 -21.6434 -21.0490 -19.6571 -18.7252 -18.1364 -18.3018
15-36 Solution We are to perform a grid independence test on the 90o diffuser (sudden expansion) case of the previous problem by refining the grid resolution. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s. Analysis We run several levels of grid refinement, tabulate the results, and plot the results. We have achieved grid independence to the third significant digit of precision by the fourth refinement level, namely, 8100 cells. The final value of ΔP is reported to three significant digits as -18.4 Pa.
-18.0
cell count 996 2700 4500 8100 13500 17855
ΔP (Pa) -19.3374 -18.5233 -18.4956 -18.4171 -18.3977 -18.3984
-18.5 ΔP (Pa) -19.0
-19.5
Discussion Newer versions of FlowLab may give slightly different -20.0 results. We could try refining the grid even further, but the increase in 100 1000 10000 100000 precision would not be worth the effort. In a flow field such as this, the Cell count separation point is fixed at the sharp corner; the CFD code does not have problems identifying the separation point. Notice that even for the case of a sudden expansion (θ = 90o) there is still a pressure recovery through the diffuser (Pin is less than Pout).
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Chapter 15 Computational Fluid Dynamics 15-37 Solution We are to generate CFD solutions for several downstream tube extension lengths to see the impact of the downstream boundary. We are also to discuss the variation of Pin with Lextend.
(a) Reverse flow
Assumptions 1 The flow is steady and incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is laminar. Properties kg/m⋅s.
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003 x
Analysis (a) Only the first case at Lextend/D2 = 0.25 has reverse flow at the pressure outlet. By Lextend/D2 = 0.50 the reverse flow problems are gone. The streamlines reveal that the flow separates at the corner, forming a recirculating eddy. The eddy reattaches at approximately x/D2 = 0.50. Streamlines are shown for the first three cases in Fig. 1. The overall streamline shapes appear to be unaffected by the downstream extent of the computational domain.
(b) Separation bubble (eddy) No reverse flow
(b) Pin, P1, and ΔP = Pin – P1 are tabulated as functions of Lextend/D2. ΔP is independent of Lextend (to three significant digits of precision) by Lextend/D2 = 0.5.
Lextend/D2 0.25 0.5 0.75 1 1.25 1.5 2 2.5 3
Pin (Pa, gage) 664.924 622.068 597.788 590.532 590.036 592.523 601.075 609.822 620.434
P1 (Pa, gage) 212.156 171.15 146.831 139.464 138.86 141.24 149.576 158.733 169.248
ΔP (Pa) 452.767 450.918 450.957 451.068 451.176 451.283 451.498 451.089 451.186
x (c) Separation bubble
x FIGURE 1 Streamlines: Lextend/D2 = (a) 0.25, (b) 0.50, and (c) 0.75. 700
(c) Inlet gage pressure Pin is plotted as a function of Lextend/D2 in Fig. 2. The inlet pressure drops as Lextend/D2 increases from 0.25 to 1.25, but then starts to slowly rise as Lextend/D2 continues to increase. We explain this trend with help from the streamlines of Fig. 1. First of all, the data for Lextend/D2 = 0.25 are not reliable, since there is reverse flow at the pressure outlet. Thus, we begin our discussion at Lextend/D2 = 0.50. The separation bubble (recirculating eddy) at the inlet to the smaller diameter tube forces the flow to converge and accelerate through an effective area that is smaller than the cross-sectional area of the small tube (a vena contracta). This high speed flow leads to a low pressure region near the separation bubble. Downstream of the separation bubble, the pressure tries to increase slightly (the downstream portion of the eddy behaves like a diffuser, recovering some of the pressure loss). However, since the outlet pressure is fixed at zero gage pressure, the inlet pressure must decrease to compensate. This explains why Pin decreases when Lextend/D2 increases from 0.50 to 1.25. By Lextend/D2 = 1.25, the outlet of the computational domain is beyond the region of pressure recovery due to the eddy, and the flow begins its slow development towards fully developed laminar pipe flow. Beyond x/D2 ≈ 1.25 the pressure decreases
650
Pin,gage (Pa) 600
550 0
1 2 Lextend/ D2
3
FIGURE 2 Inlet gage pressure as a function of normalized downstream extent of the computational domain.
15-27 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics axially along the tube because of friction along the pipe wall. But since the exit pressure is atmospheric, the inlet pressure must rise accordingly. In other words, the pressure at the inlet must rise to overcome the increasing pressure drop through the tube. We expect this trend to continue, eventually becoming linear. Based on all our results taken collectively, we recommend that Shane use a value of Lextend/D2 of 2.0 or more to ensure proper simulation of this flow. If computer speed or memory is a problem, he can get away with a minimum of Lextend/D2 = 0.75 to avoid reverse flow problems at the outlet. Discussion For other geometries, fluids, or speeds, the separation bubble may extend farther than in this example. Thus is wise to extend the computational domain reasonably far downstream (several tube diameters).
15-38 Solution We are to simulate flow through a sudden contraction in a tube, using no downstream tube extension. We are to calculate the error in pressure drop and discuss. Assumptions 1 The flow is steady and incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is laminar. Properties The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003 kg/m⋅s.
Outlet
Analysis There is no reverse flow in the CFD simulation. The streamlines (shown in the figure) reveal that since there is no downstream extension, there is obviously no way to have a separation x bubble. With the outlet pressure set to a constant value, the streamlines adjust themselves such that the fluid flows through the outlet without reverse flow. Although this may appear to be a good CFD solution, it turns out to be erroneous because the actual pressure across the interface of a sudden contraction is not constant. We obtain a net pressure difference of ΔP = Pin – Pout = 519.6 Pa. Compared to the best values from the previous problem (ΔP = 451 Pa), the percentage error is about 15%. Discussion Newer versions of FlowLab may give slightly different results. The error in ΔP caused by neglecting the downstream extension is significant. This reinforces our discussion in the text about extending outlets.
15-39 Solution We are to simulate flow through a sudden contraction in a tube, using three different values of outlet pressure at the pressure outlet boundary. We are to compare the pressure drop and discuss. Assumptions
1 The flow is steady and incompressible. 2 The flow is axisymmetric. 3 The flow is laminar.
Properties
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003 kg/m⋅s.
Analysis Results for the three cases are summarized in the table for the case in which Lextend/D2 = 2.0. We see that the inlet pressure and the pressure at x = 0 rise or fall in symphony with Pout, such that the net pressure difference is the same (to more than four significant digits) in all Pout (Pa, gage) Pin (Pa, gage) P1 (Pa, gage) ΔP (Pa) cases. The results verify a statement we -50,000 -49,398.9 -49,850.4 451.500 made in Chap. 9: For incompressible flow, it is not pressure itself which is important to 0 601.075 149.576 451.498 the flow field, but rather pressure 50,000 50,601.1 50,149.6 451.496 differences. Discussion The slight differences (in the fifth digit) in ΔP are due to numerical inaccuracies in the CFD code (the residuals never go to exactly zero). Note that we are subtracting two large numbers, which inevitably leads to precision errors. If we had run the simulation with double precision arithmetic and adjusted some of the numerical parameters, we could have obtained even better agreement.
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Chapter 15 Computational Fluid Dynamics 15-40 Solution We are to generate CFD solutions for several downstream tube extension lengths to see the impact of the downstream boundary. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties kg/m⋅s.
Separation bubble
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003
Analysis (a) Four Reynolds numbers are calculated. Note that since the diameter decreases by a factor of 4 through the contraction, the average velocity in the smaller downstream tube increases by a factor of 42 = 16 compared to that in the larger upstream tube. For the laminar case of Problem 15-37, Upstream tube: 3 ρV1 D1 ( 998.2 kg/m ) ( 0.05 m/s )( 0.008 m ) Re1 = = = 398.1 0.001003 kg/m ⋅ s μ
Downstream tube:
Re 2 =
(1)
x
FIGURE 1 Streamlines for steady, incompressible, axisymmetric, turbulent flow through a sudden contraction at Lextend/D2 = 0.75.
3 ρV2 D2 ( 998.2 kg/m ) ( 0.8 m/s )( 0.002 m ) = = 1592 0.001003 kg/m ⋅ s μ
(2)
Since both of these values are smaller than 2300, the laminar flow assumption for Problem 15-37 is reasonable. For the turbulent pipe flow of the present problem, Upstream pipe:
Downstream pipe:
3 ρV1 D1 ( 998.2 kg/m ) (1.0 m/s )( 0.8 m ) Re1 = = = 796, 200 0.001003 kg/m ⋅ s μ
(3)
3 ρV2 D2 ( 998.2 kg/m ) (16.0 m/s )( 0.2 m ) Re 2 = = = 3,185, 000 0.001003 kg/m ⋅ s μ
(4)
These Reynolds numbers are clearly high enough that the flow is indeed turbulent. (b) Our CFD solutions reveal reversed flow for the smallest two cases, i.e. for Lextend/D2 = 0.25 and 0.5. For higher values of Lextend/D2 there is no reverse flow. Comparing to the results of Problem 15-37, apparently the separation bubble for turbulent flow is somewhat longer (proportionally) than that for laminar flow. This is verified by comparing the streamlines of Fig. 1 to those of Fig. 1c of Problem 15-37. (c) Pressures Pin, P1, and ΔP = Pin – P1 are tabulated as functions of Lextend/D2 in the table. Note that we use units of kPa instead of Pa here for convenience. ΔP is independent of Lextend (to three significant digits of precision) by Lextend/D2 = 0. 5.
Lextend/D2 0.25 0.5 0.75 1 1.25 1.5 2
Pin (kPa, gage) 309.31 276.188 251.176 237.423 228.954 223.423 216.361
P1 (kPa, gage) 122.121 90.3473 65.3649 51.5843 43.0877 37.4649 30.2144
ΔP (kPa) 187.189 185.841 185.811 185.839 185.866 185.958 186.147
Discussion Newer versions of FlowLab may give slightly different results. The lack of scatter in the turbulent data for ΔP beyond Lextend/D2 = 0.5 is a rather pleasant surprise; we might have expected more scatter than the laminar flow solution of Problem 15-37, since we have two additional nonlinear transport equations to solve, with their associated interactions that complicate the solution.
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Chapter 15 Computational Fluid Dynamics 15-41 Solution We are to compare the pressure drop through a sudden contraction in a round pipe with two different outlet conditions: pressure outlet and outflow. Assumptions 1 The flow is incompressible. 2 The flow is axisymmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003 kg/m⋅s.
Analysis For the previous case (pressure outlet boundary condition), the average pressure at the inlet is Pin = 251.18 kPa, the average pressure at x = 0 is P1 = 65.36 kPa, and the pressure difference is ΔP = Pin – P1 = 185.82 kPa. For the present case (outflow boundary condition), the average pressure at the inlet is Pin = 127.94 kPa, the average pressure at x = 0 is P1 = −57.87 kPa, and the pressure difference is ΔP = Pin – P1 = 185.81 kPa. While the actual values of pressure differ throughout the contraction, the pressure difference agrees to nearly five digits of precision for the two cases. Thus we conclude that the outlet boundary condition has very little effect on this flow field. For the case with the pressure outlet boundary condition, the static pressure at the outlet of the computational domain is forced to be constant (zero gage pressure in the calculations of the previous problem). In the present case however, the outflow boundary condition does not fix static pressure – rather, it forces flow variables to level off as they approach the outlet boundary. Discussion Newer versions of FlowLab may give slightly different results. With a velocity inlet and an outflow outlet, we do not fix the value of pressure at either boundary. Instead, the CFD code assigns P = 0 gage pressure at some (arbitrary) location in the flow field (the default location is the origin). Even though the inlet and outlet pressures differ significantly between the two cases, ΔP is identical to almost five significant digits of precision. This result again verifies the statement we made in Chap. 9: For incompressible flow, it is not pressure itself which is important to the flow field, but rather pressure differences.
15-30 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics 15-42 Solution We are to generate CFD solutions for several downstream duct extension lengths to see the impact of the downstream boundary.
Separation bubble
Assumptions 1 The flow is incompressible. 2 The flow is twodimensional and symmetric about the x axis. 3 The flow is turbulent, but steady in the mean. Properties kg/m⋅s.
Reverse flow
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003
Analysis (a) Our CFD solutions reveal reversed flow for the smallest three cases, i.e. for Lextend/D2 = 0.25, 0.5, and 0.75. For higher values of Lextend/D2, there is no reverse flow. Comparing to the results of Problem 15-40, the separation bubble for 2-D flow is somewhat longer than that for axisymmetric flow. This is verified by comparing the streamlines of Fig. 1 to those of Fig. 1 of Problem 15-40.
x FIGURE 1 Streamlines for steady, incompressible, twodimensional, turbulent flow through a sudden contraction at Lextend/D2 = 0.75.
(b) Pressures Pin, P1, and ΔP = Pin – P1 are tabulated as functions of Lextend/D2 in Table 1. ΔP is independent of Lextend (to three significant digits of precision) by Lextend/D2 = 0.5. TABLE 1 Inlet pressure, average pressure at x = 0, and pressure difference as functions of downstream extent of the computational domain. Data tabulated from CFD runs for steady, incompressible, two-dimensional, turbulent flow through a sudden contraction in a duct.
Lextend/D2 0.25 0.5 0.75 1 1.25 1.5 2 3 4
Pin (kPa, gage) 18.203 17.3614 15.3649 13.7764 12.9757 12.5034 11.9787 11.5309 11.3452
P1 (kPa, gage) 7.3241 6.65408 4.67913 3.09228 2.29043 1.81676 1.29006 0.845519 0.660043
ΔP (kPa) 10.8789 10.7073 10.6858 10.6841 10.6853 10.6867 10.6887 10.6853 10.6852
Discussion Newer versions of FlowLab may give slightly different results. Comparing ΔP between this problem (2-D) and Problem 15-40 (axisymmetric), we see that ΔP for the axisymmetric case is more than 17 times greater than ΔP for the 2-D case. This is because for the axisymmetric case, the area downstream of the sudden contraction is 16 times smaller than the upstream area, while for the 2-D case, the area changes by only a factor of 4. Nevertheless, the 2-D case generates a longer separation bubble.
15-31 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics 15-43 Solution We are to generate CFD solutions for several grid resolutions to test for grid independence for flow through a jog in a channel. Specifically, we are to compare streamlines and pressure difference at each level of grid resolution. Assumptions
1 The flow is incompressible and two-dimensional. 2 The flow is turbulent, but steady in the mean.
Properties
The fluid is air with ρ = 1.225 kg/m3 and μ = 1.7894 × 10-5 kg/m⋅s.
Analysis (a) Streamlines are plotted for six grid resolution cases in Fig. 1. At the very coarse grid resolution, a small separation bubble appears at the lower left corner of the jog (Fig. 1a); flow separation is not obvious at the other corners, but it appears imminent at the most downstream corner. As grid resolution improves, the separation bubble on the lower left corner grows in size (compare Figs. 1a and 1b), and another separation bubble appears downstream of the jog. With continued improvement of the grid, both separation bubbles, especially the downstream one, continue to grow in size (Fig. 1c through 1f). Meanwhile, the streamlines become more rounded near the upper right corner of the jog. By 13,600 cells (Fig. 1e), a small recirculating zone is seen in this corner. The streamline pattern settles down by about 13,600 cells. We note that simulation of flow separation, separation bubbles, and reattachment is often a very difficult task for a CFD program. In this particular problem we must use a fairly fine grid in order to resolve all the details of the flow separation.
FIGURE 1 Streamlines for steady, incompressible, twodimensional, turbulent flow through a jog in a channel at various levels of grid resolution: cell count = (a) 900, (b) 2500, (c) 5700, (d) 9100, (e) 13,600, and (f) 18,000. The origin (0,0) is marked on each figure for reference.
(0,0)
(0,0)
(a)
(b)
(0,0)
(0,0)
(c)
(d)
(0,0)
(0,0)
(e)
(f)
(b) ΔP is tabulated as a function of cell count in the table. As grid resolution improves, ΔP increases, reflecting the effect of the larger separation bubbles. Physically, we achieve less and less pressure recovery as the separation bubbles grow. We see that ΔP becomes independent of grid resolution to four significant digits by the fifth level of resolution, namely, at 13,600 cells. This is also seen in Fig. 2 where ΔP is plotted as a function of number of cells.
cell count 900 2500 5700 9100 13600 15200 18000 20352 24840
ΔP (Pa) 2.26846 2.63636 2.71292 2.71739 2.71835 2.71828 2.71856 2.71746 2.71544
Newer versions of FlowLab Discussion may give slightly different results. The sharp corners in the streamlines of the very coarse grid resolution case (Fig. 1a) are due to interpolation errors when the CFD code calculates contours of constant stream function. The outlet of the computational domain is reasonably far downstream (ten channel heights) to avoid reverse flow at the outlet.
2.8 2.7
ΔP 2.6 (Pa) 2.5 2.4 2.3 2.2
102 100
103 104 100000 105 1000 10000 Number of cells
FIGURE 2 Pressure difference vs. number of cells.
15-32 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics 15-44 Solution We are to generate CFD solutions for several grid resolutions to test for grid independence for flow through a jog in a channel. Specifically, we are to compare streamlines and pressure difference at each level of grid resolution. Assumptions
1 The flow is incompressible. 2 The flow is two-dimensional. 3 The flow is laminar and steady.
Properties
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003 kg/m⋅s.
Analysis (a) Streamlines are plotted for six grid resolution cases in Fig. 1.
FIGURE 1 Streamlines for steady, incompressible, twodimensional, laminar flow through a jog in a channel at various levels of grid resolution: cell count = (a) 272, (b) 576, (c) 847, (d) 1700, (e) 3388, and (f) 14,123. The origin (0,0) is marked on each figure for reference. The inlet velocity for these runs is 0.10 m/s.
(0,0)
(0,0)
(a)
(b)
(0,0)
(0,0)
(c)
(d)
(0,0)
(0,0)
(e)
(f)
At the very coarse grid resolution (Fig. 1a), the streamlines are not very smooth, but reveal the overall flow pattern. Although there are no closed streamlines, it appears that flow separation is imminent at the lower left and upper right corners of the jog. The streamlines appear fairly similar to those of the turbulent flow of the previous problem. At the grid resolution of Fig. 1b, the flow clearly separates at the sharp corner at the origin. The flow also separates at the downstream inside corner of the jog. The streamlines are not well-enough resolved to show closed separation bubbles. As the grid is further refines, the separation bubbles grow in size (compare Figs. 1b through 1d). In addition, the streamlines are smoother and more rounded. With continued improvement of the grid, the streamlines become more rounded near the upper right corner of the jog. By Fig. 1e, a recirculating flow pattern is seen downstream of the jog. Based on streamline patterns, the grid is fully resolved at a cell count of about 10,000. However, the pressure drop seems to be creeping up a bit as grid resolution is refined even further; it is not clear why. (b) ΔP is tabulated as a function of grid resolution in Table 1. ΔP is also plotted as a function of number of cells in Fig. 2. As grid resolution improves, ΔP rises at first, reflecting the effect of the larger separation bubbles, but then decreases with further grid refinement, eventually leveling off after about 10,000 cells, implying grid independence by 104 cells.
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Chapter 15 Computational Fluid Dynamics Discussion Newer versions of FlowLab may give slightly different results. All cases converge nicely – the flow at this Reynolds number is steady and laminar. There is no sign of instability in the flow. The outlet of the computational domain is reasonably far downstream (ten channel heights) to avoid reverse flow at the outlet.
TABLE 1 Pressure difference from inlet to outlet as a function of cell count, V = 0.10 m/s. cell count 272 576 847 1700 3388 10625 14123 18645 23273
ΔP (Pa) 37.6118 37.8096 37.5031 36.9068 36.5925 36.4095 36.4187 36.435 36.4636
38.0
37.5 ΔP (Pa) 37.0
36.5
36.0
102 100
103 10000 104 100000 105 1000 Number of cells
FIGURE 2 Pressure difference vs. number of cells, V = 0.10 m/s.
15-34 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics 15-45 Solution We are to generate CFD solutions for several grid resolutions to test for grid independence for flow through a jog in a channel. Specifically, we are to compare streamlines and pressure difference at each level of grid resolution.
TABLE 1 Pressure difference from inlet to outlet as a function of cell count for laminar flow through a two-dimensional jog, V = 1.0 m/s.
Assumptions 1 The flow is incompressible. 2 The flow is twodimensional. 3 The flow is laminar and steady. Properties kg/m⋅s.
cell count 272 576 847 1700 3388 10625 14123 18645 23273
The fluid is water with ρ = 998.2 kg/m3 and μ = 0.001003
Analysis (a) Streamlines are plotted for six grid resolution cases in Fig. 1.
FIGURE 1 Streamlines at various levels of grid resolution: cell count = (a) 272, (b) 576, (c) 847, (d) 1700, (e) 3388, and (f) 14,123. The origin (0,0) is marked on each figure for reference. The inlet velocity for these runs is 1.0 m/s. For cases (e) and (f), the streamline patterns are at an arbitrary point in the solution. For these two cases, the CFD calculations do not converge, and eddies appear in the channel downstream of the jog. As the iterations continue, these eddies are swept downstream, but new ones appear, implying that the flow at this Reynolds number is not steady. The streamline patterns change as the CFD code iterates.
(0,0)
(0,0)
(a)
(b)
(0,0)
(0,0)
(c)
(d)
(0,0)
Flow is not converging
(e)
(0,0)
ΔP (Pa) 2025.69 2269.26 2324.99 2482.44 2582.17 2515.71 2484.86 2468.57 2511.98
Flow is not converging
(f)
15-35 PROPRIETARY MATERIAL. © 2006 The McGraw-Hill Companies, Inc. Limited distribution permitted only to teachers and educators for course preparation. If you are a student using this Manual, you are using it without permission.
Chapter 15 Computational Fluid Dynamics At the very coarse grid resolution, a small separation bubble appears at the lower left corner of the jog (Fig. 1a); flow separation is not obvious at the other corners, but it appears imminent at the most downstream corner. The streamlines appear quite similar to those of the turbulent flow of Problem 15-43. As grid resolution improves, the separation bubble on the lower left corner grows in size (compare Figs. 1a and 1b), the streamlines are smoother and more rounded, and another separation 2600 bubble appears downstream of the jog. With continued improvement of the grid, both separation bubbles continue to grow in size (Figs. 1c and 2500 1d). Meanwhile, the streamlines become more rounded near the upper right corner of the jog. By Fig. 1c, a small recirculating zone is seen in ΔP 2400 this corner. The streamline pattern continues to change somewhat as the (Pa) grid is further refined. However, just as we appear to be approaching grid 2300 independence, the flow develops some instabilities and cannot converge to a steady-state solution. This is first seen in the lower right portion of 2200 Fig. 1d – the streamlines drift away from the lower wall. While the streamline pattern in the front portion of the jog do not change 2100 appreciably, eddies develop in the recovery zone beyond the jog and are swept downstream as the grid is further refined (Figs. 1e and 1f). At the 2000 two finest grid resolutions, the flow is attempting to become unsteady. 102 103 104 100000 105 100 1000 10000 But since we are iterating towards a steady solution, the development and motion of these eddies is not physical. The solution fluctuates and is Number of cells unable to converge. The bottom line is that we really should simulate this FIGURE 2 flow with an unsteady CFD solver. Grid independence is not achieved Pressure difference as a function of the because the flow becomes unsteady and unstable when a very high number of cells in the computational resolution grid is used. domain; laminar flow through a jog in a channel. V = 1.0 m/s. (b) ΔP is tabulated as a function of grid resolution in Table 1. ΔP is also plotted as a function of number of cells in Fig. 2. As grid resolution improves, ΔP does not level off because the flow field is unstable – we should use an unsteady solver instead of a steady solver for this flow. The Reynolds number for the present problem is Re at V = 1.0 m/s:
Re =
3 ρVD ( 998.2 kg/m ) (1.0 m/s )( 0.001 m ) = = 995 0.001003 kg/m ⋅ s μ
(1)
and for the previous problem, it is 3 ρVD ( 998.2 kg/m ) ( 0.10 m/s )( 0.001 m ) Re at V = 0.10 m/s: Re = = = 99.5 0.001003 kg/m ⋅ s μ
(2)
Since the Reynolds number of the present case is ten times larger than that of the previous case, and since it is almost 1000, it is not surprising that unsteadiness and instabilities in the flow start to develop at high grid resolutions. Discussion Newer versions of FlowLab may give slightly different results. The results here are similar to those of CFD simulations of flow around a circular cylinder. Namely, the flow is naturally unsteady, and leads to convergence difficulties at high grid resolution when the simulation is forced to be steady.
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Chapter 15 Computational Fluid Dynamics 15-46 Solution We are to generate CFD solutions for compressible flow of air through a converging-diverging nozzle, and compare mass flow rate at various values of back pressure. Assumptions
1 The flow is steady and compressible. 2 The flow is axisymmetric. 3 The flow is approximated as inviscid.
Properties
The fluid is air with k = 1.4.
Analysis
We tabulate and plot m as a function of Pb/P0,inlet in Table 1 and Fig. 1 respectively.
TABLE 1 Pressure difference from inlet to outlet as a function of cell count for laminar flow through a two-dimensional jog, V = 1.0 m/s. Pb/P0,inlet 0.454545 0.545455 0.636364 0.727273 0.818182 0.909091 0.954545 0.977273 0.988636 0.990909 0.995455
9 8 7 6
m (kg/s) 5 4 3 2 1 0
m (kg/s) 9.08358 9.08356 9.08355 9.08357 9.08356 9.08354 9.03556 6.91699 4.97062 4.45686 3.16752
0
0.2
0.4
0.6
0.8
1
Pb/P0,inlet FIGURE 1 Mass flow rate as a function of back pressure ratio in a converging-diverging nozzle.
We see that m is constant as long as the flow through the throat is sonic, because the flow is choked (Pb/P0,inlet < 0.95). For values of Pb/P0,inlet around 0.95 or higher, however, the flow in the diverging section becomes subsonic (no shock waves, and not choked), and the mass flow rate decreases with increasing Pb/P0,inlet from there on, as expected, since the flow is subsonic. Discussion The small variations in m in Table 1 (in the sixth digit of precision) are due to the fact that the residuals in the CFD solution do not go to zero. A finer grid and longer run times would correct this.
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Chapter 15 Computational Fluid Dynamics 15-47 Solution We are to generate CFD solutions for compressible flow of air through a converging-diverging nozzle, and plot pressure and Mach number contours at two values of back pressure. Assumptions
1 The flow is steady and compressible. 2 The flow is axisymmetric. 3 The flow is approximated as inviscid.
Properties
The fluid is air with k = 1.4.
Analysis We plot contours of P and Ma in Fig. 1. For the case in which Pb/P0,inlet = 0.455 (Fig. 1a), we observe a shock wave in the diverging portion of the nozzle. When Pb/P0,inlet = 0.977 (Fig. 1b), however, the flow in the entire nozzle is subsonic, and there are no shock waves.
P
(a) FIGURE 1 Contour plots of P and Ma for compressible flow of air through a converging-diverging nozzle: Pb/P0,inlet = (a) 0.455, and (b) 0.977. A normal shock is seen for the first case, since the flow is choked, but no shock is seen in the second case, since the flow is subsonic everywhere. The pressure scale is 0 (blue) to 220 kPa (red) for both cases. The Mach number scale is 0 (blue) to 2.5 (red) for both cases.
Ma
P
(b) Ma
Plots of P and Ma versus x are shown in Fig. 2 for the two cases. Also shown are calculations from one-dimensional inviscid theory. The agreement is excellent. Discussion Newer versions of FlowLab may give slightly different results. The shock wave calculated by CFD is not straight, but curved, since the calculations are axisymmetric, and therefore show more detail than the simplified onedimensional approximation.
15-48 Solution
We are to repeat the previous problem, but for the axisymmetric case.
Analysis The results are expected to be similar, except that for the 2-D case, the area change is much less significant. Therefore it the value of back pressure that causes the flow to choke will be different than the axisymmetric case. We should still be able to observe a normal shock in the flow at low-enough values of back pressure. Discussion
Newer versions of FlowLab may give slightly different results.
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Chapter 15 Computational Fluid Dynamics 15-49 Solution
We are to study the effect of rear-end shape on automobile drag coefficient.
Analysis The template provides five different geometries. Model 1 has a very blunt rear end, kind of like a station wagon. As the model number increases, the rear end TABLE 1 gets more slanted and rounded. Table 1 shows the Drag coefficient as a function of model number, where the calculated drag coefficient for each model. rear end of the car is modified according to model number. It turns out that Model number 4 has the lowest Model Description CD drag coefficient, and Model number 1 has the number highest. We probably would have guessed the latter, 1 Blunt rear end, like a station wagon 0.320 but not the former. Namely, the car with the most Back window with a short trunk blunt rear end (Model 1) has the highest drag as we 2 section 0.298 might expect, but most people would predict that Similar to model 2, but with a longer Model 5, which is the most rounded, would have the 3 trunk 0.276 lowest drag. Model 4 has a short notch for the trunk, More rounded back end with a short 0.180 and the aerodynamics turn out such that it has the 4 notch for the trunk lowest drag. 5 Fully rounded back end 0.212 Streamline plots are shown in Fig. 1 for Models 1 and 4, the highest and lowest drag cases, respectively. It is clear from these plots that the blunted rear body has a larger separation bubble in the wake (low pressure in the wake which leads to large drag). On the other hand, Model 4 has a much smaller wake and therefore much less drag. Model
(a)
1
(b)
4
FIGURE 1 Streamlines for two representative twodimensional automobile shapes: (a) Model 1, and Model 4.
Discussion Newer versions of FlowLab may give slightly different results. It is not always immediately obvious whether a shape will have more or less drag than another shape. These results are two-dimensional, whereas actual automobiles, of course, are three-dimensional [see Problem 15-52].
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Chapter 15 Computational Fluid Dynamics 15-50 Solution
We are to examine the effect of the location of the upstream boundary on automobile drag calculations.
Analysis We run five cases, and show the results in Table 1. It turns out that by about 20 heights away, the upper boundary condition no longer impacts the solution significantly to within three significant digits. The final drag coefficient is reported as 0.180 to three significant digits.
These data are also plotted in Fig. 1. We use a log scale on the horizontal axis since the range of H/h is fairly large. For three digits of precision, it is necessary that the top boundary be at least 20 car heights tall. If H/h is shorter than this, the upper boundary of the computational domain adversely influences the flow field. In the real-life flow, of course, the upper boundary is nearly 0.22 infinite. 0.21
TABLE 1 Drag coefficient as a function of the normalized extent of the computational domain for turbulent flow over a cylindrical block. H/h 5 10 20 35 50 100
0.2
CD 0.19
CD 0.2101 0.1847 0.1798 0.1796 0.1797 0.1798
0.18 0.17 0 10 1
Discussion Newer versions of FlowLab may give slightly different results. The bottom wall of the computational domain is not moving in these calculations, so the flow near the ground is not modeled properly. Furthermore, the calculations are two-dimensional, while a real car is of course fully three-dimensional.
15-51 Solution
1 10 10 h/H
102 100
FIGURE 1 Drag coefficient plotted as a function of the normalized extent of the computational domain for turbulent flow over a twodimensional automobile.
We are to compare turbulence models for the calculation of automobile drag.
Analysis The results from the CFD calculations are presented in Table 1. There is some variation in the calculated values of CD depending on the turbulence model used; CD ranges from 0.175 for the Spallart-Allmaras model to 0.223 for the RSM model. The range of scatter is about 12%, which is actually not that large for comparison of four very different turbulence models. It is impossible to say which one, if any, is correct, since all turbulence models are approximations, with calibrated constants. Furthermore, we have no experimental data with which to compare the CFD results.
TABLE 1 Predicted drag coefficient on a two-dimensional automobile as a function of turbulence model. Turbulence model Spallart-Allmaras (1 eq.) k-ε (2 eq.) k-ω (2 eq.) Reynolds stress model (7 eq.)
CD 0.175 0.182 0.221 0.223
Discussion Newer versions of FlowLab may give slightly different results. It would be good if we had experimental data with which to compare.
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Chapter 15 Computational Fluid Dynamics 15-52 Solution
We are to compare 2-D and 3-D drag predictions for flow over an automobile.
Analysis The 3-D drag coefficient is 0.785, significantly higher than the 2-D case, which is around 0.2 for most of the cases. This is most likely due to the fact that flow separates off the sides of the car as well as off the top. Note too that the car model used in this analysis is more like a truncated 2-D model rather than a truly three-dimensional model. If all the sharp corners were rounded off, the drag coefficient would be much lower than that calculated here. Discussion Newer versions of FlowLab may give slightly different results. This is not a very good drag coefficient for a modern car, and there is much improvement possible by further streamlining.
15-53 Solution
We are to examine entrance the length for laminar pipe flow.
Analysis Table 1 shows the entrance length calculations, along with a comparison with theory. The results are very good, as is expected for laminar pipe flow. Note that the results are somewhat subjective since the end of the developing region is being judged “by eye”. Figure 1 shows the pressure along the axis as a function of downstream distance along the pipe axis for the case in which Re = 500. It is clear that the pressure drop is more severe (higher slope) at the beginning section of the pipe, but it is hard to tell at what axial location the flow becomes fully developed from this type of a plot. Thus, the entrance length is more appropriately determined by
TABLE 1 Entry length vs. Reynolds number for developing laminar pipe flow. Re 500 1000 1500 2000
Le/D, CFD 20 50 86 1050
Le/D, Theory 30 60 90 120
studying the velocity profiles, as above.
FIGURE 1 Pressure as a function of axial distance down a long straight pipe at Re = 500.
0.0007 0.0006 0.0005 P (Pa) 0.0004 0.0003 0.0002 0.0001 0 0
50
100 x (m)
150
200
Discussion Newer versions of FlowLab may give slightly different results. Also, student answers may vary considerably because of the subjectivity of these results, as mentioned above. There are no shed vortices, unsteadiness, or non-symmetries in straight laminar pipe flow, so it is no surprise that the CFD calculations perform so well.
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Chapter 15 Computational Fluid Dynamics 15-54 Solution flow.
We are to study the entrance region in turbulent pipe
Analysis Table 1 shows the entrance length calculations, along with a comparison with the empirical equation. The CFD results are consistently low compared to the empirical results, although the agreement improves as Reynolds number increases. Since the grid is sufficiently resolved and y+ values are appropriate, the discrepancies are due to deficiencies in the turbulence model. Discussion Newer versions of FlowLab may give slightly different results.
15-55 Solution with theory.
TABLE 1 Entry length vs. Reynolds number for developing turbulent pipe flow. Re 2000 5000 10000 20000 50000
Le/D, CFD 10 11 13 18 22
Le/D, Exper. 15.6 18.2 20.4 22.9 26.7
We are to study fully developed laminar pipe flow, and compare CFD calculations of Darcy friction factor
Analysis Table 1 shows the CFD calculations of f, along with a comparison with theory. The CFD and theoretical results agree very well, since that the grid is well-enough resolved, and the flow is laminar. The values agree to within 0.05%, which is as good as one can ever expect from a CFD analysis. TABLE 1 Darcy friction factor vs. Reynolds number for fully developed laminar pipe flow – comparison between CFD calculations and theory (f = 64/Re). Re 500 750 1000 1250 1500 2000
f, CFD 0.127944 0.0852963 0.0639721 0.0511774 0.0426476 0.0319855
f, Theory 0.128 0.085333 0.064 0.0512 0.042667 0.032
Error (%) -0.04375 -0.0434 -0.04359 -0.04414 -0.04469 -0.04531
Discussion Newer versions of FlowLab may give slightly different results. There are no shed vortices, unsteadiness, or non-symmetries in straight laminar pipe flow, so it is no surprise that the CFD calculations perform so well.
15-56 Solution We are to study fully developed turbulent pipe flow, with smooth walls and compare CFD calculations of Darcy friction factor with experimentally determined values.
TABLE 1 Darcy friction factor vs. Reynolds number for fully developed turbulent pipe flow.
Analysis Table 1 shows the CFD calculations of f, along with a Re comparison with the empirical formula (the Colebrook equation). The 5000 CFD and theoretical results agree reasonably well (within less than 10000 10% for all cases tested), and the agreement improves with increasing 50000 Reynolds number. The k-ε model performs better at the higher values of Reynolds number because the turbulence model is calibrated for high Re flows.
f, CFD 0.040007 0.03293 0.022029
f, Exper. 0.0374 0.0309 0.0209
Discussion Newer versions of FlowLab may give slightly different results. Ten percent agreement is excellent, considering that the Colebrook equation (or the Moody chart) is accurate to only about 15% to begin with.
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Chapter 15 Computational Fluid Dynamics 15-57 Solution We are to study fully developed turbulent pipe flow with rough walls, and compare CFD calculations of Darcy friction factor with experimentally determined values.
TABLE 1 Darcy friction factor vs. dimensionless roughness height for fully developed turbulent pipe flow at Re = 1 × 106.
Analysis Table 1 shows the CFD calculations of f, along with a comparison with the empirical formula (the Colebrook equation). The CFD and theoretical results agree reasonably well (within less than 15% for all cases tested), and the agreement improves with increasing roughness. Since the agreement is within the known inaccuracy of the Colebrook equation (about 15%), these CFD results are considered adequate.
ε/D 0.00005 0.00025 0.0005 0.001 0.0015 0.002
Discussion Newer versions of FlowLab may give slightly different results.
15-58 Solution
f, CFD 0.0113382 0.0134436 0.016288 0.0204516 0.0231213 0.0246173
f, Exper. 0.0126 0.0152 0.0172 0.0199 0.0220 0.0236
We are to model the laminar boundary layer on a flat plate using CFD, and compare to analytical results.
Analysis For a Reynolds number of 1 × 105, the CFD calculations give a nondimensional boundary layer thickness of δ /x = 0.0154 and a drag coefficient of CD = 0.00435. The theoretical values are obtained from equations in Chaps. 10 and 11, namely,
δ x
=
4.91 Re x
=
4.91 1× 105
= 0.0155
(1)
= 0.00421
(2)
and CD = C f =
1.33 Re x
=
1.33 1× 105
The agreement is excellent for both values, the discrepancy being less than 1% for δ/x and about 3% for drag coefficient. Discussion
Newer versions of FlowLab may give slightly different results.
15-59 Solution
We are to compare CFD results to experimental results for the case of a flat plate turbulent boundary layer.
Analysis For a Reynolds number of 1 × 107, the CFD calculations give a nondimensional boundary layer thickness of δ/x = 0.0140 and a drag coefficient of CD = 0.00292. The empirical values are obtained from equations in Chaps. 10 and 11, namely,
δ x
=
0.38 0.38 = = 0.015 1/ 5 1/ 5 Re x 1× 107
(
)
and
CD = C f =
0.074 0.074 = = 0.0029 1/ 5 1/ 5 Re x 1× 107
(
)
The agreement is excellent for both values, the discrepancy being about 7% for δ/x and negligible (within 2 significant digits) for drag coefficient. Discussion Newer versions of FlowLab may give slightly different results. We report our empirical values to only two significant digits in keeping with the level of precision and accuracy of turbulent flows.
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Chapter 15 Computational Fluid Dynamics 15-60 Solution
We are to compare turbulence models in CFD calculations of a flat plate turbulent boundary layer.
Analysis The calculations for drag coefficient and boundary layer thickness are shown in Table 1 for four turbulence models: standard k-ε, standard k-ω, Spallart-Allmaras, and the TABLE 1 Reynolds stress model (RSM). The first two are two-equation models, Drag coefficient and boundary layer the third is a one-equation model, and the fourth is a full Reynolds thickness for a flat plate at Re = 1 × 107 , as stress model (5 equations for the two-dimensional case). All predicted by CFD with various turbulence turbulence models are approximations, with calibrated constants. The models. drag coefficients range from 0.0027 for the RSM model to 0.0029 for the k-ε model. The empirical value is 0.0029. The dimensionless CD ε/D δ /L boundary layer thickness ranges from 0.012 for the k-ω model to k-ε 0.00291589 0.0140152 0.014 for the k-ε model. The empirical value is 0.015. Thus, all the k-ω 0.00281271 0.0122085 turbulence models do very well at predicting this flow, and the k-ε S-A 0.00276828 0.013329 model performs the best, overall. The RSM model, in spite of its RSM 0.00274706 0.0131455 increased complexity, does not do as well as some of the simpler models. Since all turbulence models are calibrated with the turbulent flat plate boundary layer, it is not surprising that all of them give reasonable results. Discussion
Newer versions of FlowLab may give slightly different results.
15-61 Solution
We are to compare the velocity and thermal boundary layer thicknesses on a heated flat plate, laminar flow.
Analysis The results are shown in Table 1. We denote the thermal boundary layer thickness as δT /L, and compare to the velocity boundary layer thickness, which we denote as δu /L. The TABLE 1 velocity (or momentum) boundary layer thicknesses are identical, as Nondimensional temperature and boundary expected, since the Reynolds number is the same for either fluid. layer thicknesses for a heated flat plate at However, since the Prandtl number (Pr = ν/κ) of water is close to Re = 1 × 105 , as predicted by CFD (laminar 1000, which is much greater than one, the velocity boundary layer flow). thickness is much greater than the thermal boundary layer thickness for Fluid δT /L δu /L the water flow. In other words, momentum diffuses into the freeair 0.017313 0.015412 stream flow much more rapidly than does temperature since the water 0.007664 0.015412 Prandtl number is large. The two thicknesses are nearly equal for the air, although the thermal boundary layer thickness is somewhat higher than the momentum boundary layer thickness. This is to be expected since the Prandtl number in air is about 0.70, which means that temperature diffuses more rapidly than does momentum since Pr < 1. Discussion Newer versions of FlowLab may give slightly different results. Since the flow is laminar and the grid is well resolved with a large computational domain, these results are nearly “exact”.
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Chapter 15 Computational Fluid Dynamics 15-62 Solution flow.
We are to compare the velocity and thermal boundary layer thicknesses on a heated flat plate, turbulent
TABLE 1 Analysis The results are shown in Table 1. We denote the Nondimensional temperature and boundary thermal boundary layer thickness as δT /L, and compare to the velocity layer thicknesses for a heated flat plate at boundary layer thickness, which we denote as δu /L. The velocity (or Re = 1 × 107 , as predicted by CFD momentum) boundary layer thicknesses are identical, as expected, (turbulent flow, using the standard k-ε since the Reynolds number is the same for either fluid. The velocity turbulence model). boundary layer thickness is slightly greater than the thermal boundary Fluid δT /L δu /L layer thickness for the water flow, while the two thicknesses are nearly air 0.014527 0.014015 equal for the air, although the thermal boundary layer thickness is water 0.013162 0.014014 slightly greater than the momentum boundary layer thickness. These small differences are attributed to differences in the Prandtl number, as explained for the laminar case of the previous problem. However, turbulent diffusion dominates over laminar diffusion, and therefore the effect of Prandtl number is minimal. The thermal and momentum boundary layer thicknesses are nearly equal in both fluids since turbulent diffusion effects dominate laminar (molecular) diffusion effects, especially at high Re. Discussion Newer versions of FlowLab may give slightly different results. Mass (species), momentum (velocity), and energy (temperature) diffuse nearly equally in a turbulent flow since the large turbulent eddies cause rapid mixing through the boundary layer. Turbulence models do not actually calculate the details of the unsteady flow caused by these turbulent eddies; rather, they model the increased diffusion effects with approximations that enable the calculations to be performed in reasonable time on a computer.
15-63 Solution
We are to calculate and plot pressure drop down a pipe with an elbow in turbulent flow.
Analysis A plot of pressure as a function of axial distance along the upstream pipe, through the elbow, and through the downstream pipe is shown in Fig. 1. The pressure losses are nearly linear through the entrance region. In the elbow itself, and just downstream of it, the pressure drops rapidly. The rate of pressure drop in the pipe section downstream of the elbow is about the same as that upstream. Therefore, it appears that most of the pressure drop occurs near the region of the elbow. 0.35
Upstream pipe section
0.3 0.25
FIGURE 1 Pressure along the axis of a pipe with an elbow at Re = 2 × 104, as predicted by CFD (turbulent flow, using the standard k-ε turbulence model).
Discussion
Elbow
P, Pa 0.2 0.15
Downstream pipe section
0.1 0.05 0 0
10 20 30 40 Axial distance x along the pipe & elbow, m
50
Newer versions of FlowLab may give slightly different results.
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Chapter 15 Computational Fluid Dynamics 15-64 Solution
We are to study the counter-rotating vortices in a turbulent pipe flow downstream of an elbow.
Analysis Velocity vectors at several cross sections are plotted in Fig. 1. There are no counter-rotating eddies upstream of the elbow. However, they are formed as the fluid passes through the elbow, and are very strong just downstream of the elbow. These vortices decay in strength down the pipe after the elbow, but they persist for a very long time, and may influence the accuracy of flow meters downstream of an elbow. This is why many manufacturers of pipe flow meters recommend that their flow meter be installed at least 10 or 20 pipe diameters downstream of an elbow – to avoid influence of the counter-rotating eddies. Downstream pipe section
FIGURE 1 Velocity vector plots at several cross sections of a pipe with an elbow at Re = 2 × 104, as predicted by CFD (turbulent flow, using the standard k-ε turbulence model).
Elbow
Upstream pipe section
Flow direction
Discussion
The counter-rotating eddies lead to additional irreversible head loss as they dissipate.
15-65 Solution
We are to calculate the minor loss coefficient through a pipe elbow in turbulent flow.
Analysis The value of KL given in Chap. 8 is 0.30. For the pipe with the elbow, the pressure drop calculated by the CFD code is 0.284 kPa. For the straight pipe, the pressure drop is calculated to be 0.224 kPa. Subtracting these and converting to minor loss coefficient, the value of KL predicted by our CFD calculation for the standard k-ε turbulence model is 0.295 – a difference of less than two percent, and well within the accuracy of the Colebrook equation, which is about 15%, and the accuracy of the tabulated minor loss coefficients, which is often much greater than 15%. This agreement is better than expected, considering that this is a very complex 3-D flow, and turbulence models may not necessarily apply for such problems. Discussion The agreement here is excellent – CFD does not always match so well with experiment, especially in turbulent flow when using turbulence models, since turbulence models are approximations that often lead to significant error.
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Chapter 15 Computational Fluid Dynamics 15-66 Solution
We are to compare various turbulence models – how well they predict the minor loss in a pipe elbow.
Analysis The results for four turbulence models are listed in Table 1. The standard k-ε turbulence model does the best job in predicting KL. The standard k-ω turbulence model does the worst job. The Spallart-Allmaras turbulence model calculations are not the worst, even though this is the simplest of the models. One would hope TABLE 1 Minor loss coefficient as a function of turbulence model for flow that the more complicated model (Reynolds stress through a 90o elbow in a pipe. The error is in comparison to the model) would do a better job than the simpler experimental value of 0.30. models, but this is not the case in the present problem (it does worse than the k-ε model, but Turbulence model KL Error (%) better than the other two. All turbulence models Spallart-Allmaras (1 eq.) 0.204 -32% are approximations, with calibrated constants. -1.7% kε (2 eq.) 0.295 While one model may do a better job in a certain 34% k-ω (2 eq.) 0.401 flow, it may not do such a good job in another Reynolds stress model (7 eq.) 0.338 13% flow. This is the unfortunate state of affairs concerning turbulence models. Discussion Newer versions of FlowLab may give slightly different results. Although the RSM model does not seem to be too impressive based on this comparison, keep in mind that this is a very simple flow field. There are flows (generally flows of very complex geometries and rotating flows) for which the RSM model does a much better job than any 1- or 2equation turbulence model, and is worth the required increase in computer resources.
15-67 Solution We are to use CFD to calculate the lift and drag coefficients on an airfoil as a function of angle of attack. Analysis The CFD analysis involves turbulent flow, using the standard k-ε turbulence model. The results are tabulated and plotted. The lift coefficient rises to 1.44 at α = 14o, beyond which the lift coefficient drops off. So, the stall angle is about 14o. Meanwhile, the drag coefficient increases slowly up to the stall location, and then rises significantly after stall. 1.6 1.4 C L 1.2 & 1 C D 0.8 0.6 0.4 0.2 0
CL
α (degrees) -2 0 2 4 5 6 8 10 12 14 16 18 20
CL 0.138008 0.348498 0.560806 0.769169 0.867956 0.967494 1.14544 1.29188 1.39539 1.44135 1.41767 1.34726 1.29543
CD 0.0153666 0.0148594 0.0149519 0.0170382 0.0192945 0.0210042 0.0275433 0.0375832 0.0522318 0.0725146 0.100056 0.140424 0.274792
CD -5
0
5 10 Angle of attack α, degrees
15
20
Discussion We note that this airfoil is not symmetric, as can be verified by the fact that the lift coefficient is nonzero at zero angle of attack. The lift coefficient does not drop as dramatically as is observed empirically. Why? The flow becomes unsteady for angles of attack beyond the stall angle. However, we are performing steady calculations. For higher angles, the run does not even converge; the CFD calculation is stopped because it has exceeded the maximum number of allowable iterations, not because it has converged. Thus the main reason for not capturing the sudden drop in CL after stall is because we are not accounting for the transient nature of the flow. The airfoil used in these calculations is called a ClarkY airfoil.
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Chapter 15 Computational Fluid Dynamics 15-68 Solution We are to analyze the effect of Reynolds number on lift and drag coefficient. Analysis The CFD analysis involves turbulent flow, using the standard k-ε turbulence model. For this airfoil, which is different than the airfoil analyzed in the previous problem, and for the case in which Re = 3 × 106, the lift coefficient rises to about 1.77 at about 18o, beyond which the lift coefficient drops off (see the first table). So, the stall angle is about 18o. Meanwhile, the drag coefficient increases slowly up to the stall location, and then rises significantly after stall. The data are also plotted below.
Re = 3 × 106
α (degrees)
CL 0.221797 0.65061 0.858744 1.05953 1.2501 1.42542 1.57862 1.69816 1.76686 1.75446 1.70497
0 4 6 8 10 12 14 16 18 20 22
Re = 3 × 106 2 1.5
Re = 6 × 106
CL
CL & 1 CD
α (degrees)
CD 0 5
10 15 Angle of attack α, degrees
20
25
Re = 6 × 106
2 1.5 CL & 1 CD
CL
0.5
CD 0 0
5
10 15 Angle of attack α, degrees
20
25
CL 0.226013 0.659469 0.870578 1.07512 1.27094 1.4533 1.61589 1.74939 1.83901 1.85799 1.76048
0 4 6 8 10 12 14 16 18 20 22
0.5
0
CD 0.0118975 0.0166523 0.0212052 0.0273125 0.0351061 0.0447038 0.0562746 0.0702321 0.0875881 0.111326 0.178404
CD 0.0106894 0.015384 0.0198087 0.0256978 0.0331355 0.0422083 0.0530114 0.0657999 0.0812925 0.101563 0.138806
For the case in which Re = 6 × 106, lift coefficient rises to about 1.86 at about 20o, beyond which the lift coefficient drops off (see the second table). So, the stall angle is about 20o. Meanwhile, the drag coefficient increases slowly up to the stall location, and then rises significantly after stall. These data are also plotted. 2
The maximum lift coefficient and the stall angle have both increased somewhat compared to those at Re = 3 × 106 (half the Reynolds number). Apparently, the higher Reynolds number leads to a more vigorous turbulent boundary layer that is able to resist flow separation to a greater downstream distance than for the lower Reynolds number case. For all angles of attack, the drag coefficient is slightly smaller for the higher Reynolds number case, reflecting the fact that the skin friction coefficient decreases with increasing Re along a wall, all else being equal. [Airfoil drag (before stall) is due mostly to skin friction rather than pressure drag.] Finally, we plot the lift coefficient as a function of angle of attack for the two Reynolds numbers. The airfoil clearly performs better at the higher Reynolds number.
1.8 1.6 1.4 1.2
CL
Re = 6 × 106 Re = 3 × 106
1 0.8 0.6 0.4 0.2
0 Discussion The behavior of the lift and drag coefficients beyond stall is not 0 10 20 30 α, degrees as dramatic as we might have expected. Why? The flow becomes unsteady for angles of attack beyond the stall angle. However, we are performing steady calculations. For higher angles, the run does not even converge; the CFD calculation is stopped because it has exceeded the
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Chapter 15 Computational Fluid Dynamics maximum number of allowable iterations, not because it has converged. Thus the main reason for not capturing the sudden drop in CL after stall is because we are not accounting for the transient nature of the flow. We note that the airfoil used in this problem (a NACA2415 airfoil) is different than the one used in Problem 15-67 (a ClarkY airfoil). Comparing the two, the present one has better performance (higher maximum lift coefficient and higher stall angle, even though the Reynolds numbers here are lower than that of Problem 15-67. At higher Re, this airfoil may perform even better.
15-69 Solution number.
We are to examine the effect of grid resolution on airfoil stall at a given angle of attack and Reynolds
Analysis The CFD results are shown in Table 1 for the case in which the airfoil is at a 15o angle of attack at a Reynolds number of 1 × 107. The lift coefficient levels off to a value of 1.44 to three significant digits by a cell count of about 17,000. The drag coefficient levels off to a value of 0.849 to three significant digits by a cell count of about 22,000. Thus, we have shown how far the grid must be refined in order to achieve grid independence. Thus, we have achieved grid independence for a cell count greater than about 20,000. As for the effect of grid resolution on stall angle, we see that with poor grid resolution, flow separation is not predicted accurately. Indeed, when the grid resolution is poor (under 10,000 cells in this particular case), stall is not observed even though the angle of attack (15o) is above the stall angle (14o) for this airfoil at this Reynolds number (1 × 107). When the cell count is about 15,000, however, stall is observed. Thus, yes, grid resolution does affect calculation of the stall angle – it is not predicted well unless the grid is sufficiently resolved. Discussion
TABLE 1 Lift and drag coefficients as a function of cell count (higher cell count means finer grid resolution) for the case of flow over a 2-D airfoil at an angle of attack of 15o and Re = 1 × 107. Cells 672 1344 2176 6264 12500 16800 21700 24320 27200
CL 1.09916 1.00129 1.06013 1.02186 1.42487 1.4356 1.43862 1.43719 1.43741
CD 0.242452 0.224211 0.196212 0.189023 0.0869799 0.0857791 0.0848678 0.0852304 0.0854769
Newer versions of FlowLab may give slightly different results.
15-70 Solution We are to study the effect of computational domain extent on the calculation of drag in creeping flow. R/L CD Analysis The drag coefficient is 3 334.067 listed as a function of R/L in the table. The 10 289.152 data are also plotted. From these data we 50 278.291 see that for R/L greater than about 50, the 100 277.754 drag coefficient has leveled off to a value 500 277.647 of about 278 to three significant digits. 1000 277.776 This is rather surprising since the 2000 278.226 Reynolds number is so small, and the viscous effects are expected to influence the flow for tens of body lengths away from the body. Discussion When analyzing creeping flow using CFD, it is important to extend the computational domain very far from the object of interest, since viscous effects influence the flow very far from the object. This effect is not as great at high Reynolds numbers, where the inertial terms dominate the viscous terms.
350 340 330 320 310 CD 300 290 280 270 260 250 1010
1 10 10
102 100 R/L
103 10000 104 1000
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Chapter 15 Computational Fluid Dynamics 15-71 Solution We are to study the effect of Reynolds number on flow over an ellipsoid.
TABLE 1 Drag coefficient as a function of Reynolds number for flow over a 2×1 ellipsoid.
1000
Analysis The velocity profile 100 for creeping flow (Re < 1) shows a Re CD very slowly varying velocity from 0.1 552.89 CD zero at the wall to V eventually. At 0.5 114.634 high Re, we expect a thin boundary 5 14.3342 layer and flow that accelerates 20 4.89852 10 around the body. However, in 50 2.61481 creeping flow, there is negligible 100 1.691 inertia, and the flow does not accelerate around the body. Instead, 1 the body has significant impact on the flow to distances very far from the -1 0 1 10 10 10 103 0.1 1 10 100 body. As Re increases, the drag coefficient drops sharply, as expected Re based on experimental data (see Chap. 11). At the higher values of Re FIGURE 1 (here, for Re = 50 and 100), inertial effects are becoming more Drag coefficient plotted as a function of significant than viscous effects, and the velocity flow disturbance caused Reynolds number for creeping flow over a by the body is confined more locally around the body compared to the 2×1 ellipsoid. lower Reynolds number cases. If Re were to be increased even more, very thin boundary layers would develop along the walls. The data are also plotted in Fig. 1. The drop in drag coefficient with increasing Reynolds number is quite dramatic as Re ranges from 0.1 to 100. (CD decreases from more than 500 to nearly 1 in that range). Thus, we use a log-log scale in Fig. 1. Discussion
Newer versions of FlowLab may give slightly different results.
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Chapter 15 Computational Fluid Dynamics
General CFD Problems
15-72 Solution We are to generate three different coarse grids for the same geometry and node distribution, and then compare the cell count and grid quality. Analysis The three meshes are shown in Fig. 1. The node distributions along the edges of the computational domain are identical in all three cases, and no smoothing of the mesh is performed. The structured multi-block mesh is shown in Fig. 1a. We split the domain into four blocks for convenience, and to achieve cells with minimal skewing. There are 1060 cells. The unstructured triangular mesh is shown in Fig. 1b. There is only one block, and it contains 1996 cells. The unstructured quad mesh is shown in Fig. 1c. It has 833 cells in its one block. Comparing the three meshes, the triangular unstructured mesh has too many cells. The unstructured quad mesh has the least number of cells, but the clustering of cells occurs in undesirable locations, such as at the outlets on the right. The structured quad mesh seems to be the best choice for this geometry – it has only about 27% more cells than the unstructured quad mesh, but we have much more control on the clustering of the cells. Skewness is not a problem with any of the meshes.
(a)
(b)
(c)
FIGURE 1 Comparison of three meshes: (a) structured multiblock, (b) unstructured triangular, and (c) unstructured quadrilateral. Discussion results.
Depending on the grid generation software and the specified node distribution, students will get a variety of
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Chapter 15 Computational Fluid Dynamics 15-73 Solution We are to run a laminar CFD calculation of flow through a wye, calculate the pressure drop and how the flow splits between the two branches. Analysis We choose the structured grid for our CFD calculations. The back pressure at both outlets is set to zero gage pressure, and the average pressure at the inlet is calculated to be -8.74 × 10-5 Pa. The pressure drop through the wye is thus only 8.74 × 10-5 Pa (a negligible pressure drop). The streamlines are shown in Fig. 1. For this case, 57.8% of the flow goes out the upper branch, and 42.2% goes out the lower branch. Discussion There appears to be some tendency for the flow to separate at the upper left corner of the branch, but there is no reverse flow at the outlet of either branch. This case is compared to a turbulent flow case in the following problem.
Upper branch
Lower branch
FIGURE 1 Streamlines for laminar flow through a wye.
15-74 Solution We are to run a turbulent CFD calculation of flow through a wye, calculate the pressure drop and how the flow splits between the two branches. Analysis We choose the structured grid for our CFD calculations. The back pressure at both outlets is set to zero gage pressure, and the average pressure at the inlet is calculated to be -3.295 Pa. The pressure drop through the wye is thus 3.295 Pa (a significantly higher pressure drop than that of the laminar flow, although we note that the inlet velocity for the laminar flow case was 0.002 times that of the turbulent flow case). The streamlines are shown in Fig. 1. For this case, 54.4% of the flow goes out the upper branch, and 45.6% goes out the lower branch. Compared to the laminar case, a greater percentage of the flow goes out the lower branch for the turbulent case. The streamlines at first look similar, but a closer look reveals that the spacing between streamlines in the turbulent case is more uniform, indicating that the velocity distribution is also more uniform (more “full”), as is expected for turbulent flow.
Upper branch
Lower branch
FIGURE 1 Streamlines for turbulent flow through a wye. The k-ε turbulence model is used.
Discussion There appears to be some tendency for the flow to separate at the upper left corner of the branch, but there is no reverse flow at the outlet of either branch.
15-75 Solution
We are to keep refining a grid until it becomes grid independent for the case of a laminar boundary layer.
Analysis Students will have varied results, depending on the grid generation code, CFD code, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-76 Solution
We are to keep refining a grid until it becomes grid independent for the case of a turbulent boundary layer.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
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Chapter 15 Computational Fluid Dynamics 15-77 Solution
We are to study ventilation in a simple 2-D room using CFD, and using a structured rectangular grid.
Analysis Students will have varied results, depending on the grid generation code, CFD code, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-78 Solution
We are to repeat the previous problem except use an unstructured grid, and we are to compare results.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-79 Solution
We are to use CFD to analyze the effect of moving the supply and/or return vents in a room.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-80 Solution
We are to use CFD to analyze a simple 2-D room with air conditioning and heat transfer.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-81 Solution
We are to compare the CFD predictions for 2-D and 3-D ventilation.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
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Chapter 15 Computational Fluid Dynamics 15-82 Solution We are to use CFD to study compressible flow through a converging nozzle with inviscid walls. Specifically, we are to vary the pressure until we have choked flow conditions. Analysis Students will have varied results, depending on the grid generation code, CFD code, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-83 Solution We are to repeat the previous problem, but allow friction at the wall, and also use a turbulence model. We are then to compare the results to those of the previous problem to see the effect of wall friction and turbulence on the flow. Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-84 Solution
We are to generate a low-drag, streamlined, 2-D body, and try to get the smallest drag in laminar flow.
Analysis Students will have varied results, depending on the grid generation code, CFD code, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-85 Solution We are to generate a low-drag, streamlined, axisymmetric body, and try to get the smallest drag in laminar flow. We are also to compare the axisymmetric case to the 2-D case of the previous problem. Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-86 Solution We are to generate a low-drag, streamlined, axisymmetric body, and try to get the smallest drag in turbulent flow. We are also to compare the turbulent drag coefficient to the laminar drag coefficient of the previous problem. Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
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Chapter 15 Computational Fluid Dynamics 15-87 Solution We are to use CFD to study Mach waves in supersonic flow. We are also to compare the computed Mach angle with that predicted by theory. Analysis Students will have varied results, depending on the grid generation code, CFD code, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-88 Solution theory.
We are to study the effect of Mach number on the Mach angle in supersonic flow, and we are to compare to
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
Review Problems
15-89C Solution (a) False: If the boundary conditions are not correct, if the computational domain is not large enough, etc., the solution can be erroneous and nonphysical no matter how fine the grid. (b) True: Each component of the Navier-Stokes equation is a transport equation. (c) True: The four-sided cells of a 2-D structured grid require less cells than do the triangular cells of a 2-D unstructured grid. (Note however, that some unstructured cells can be four-sided as well as three-sided.) (d) True: Turbulence models are approximations of the physics of a turbulent flow, and unfortunately are not universal in their application.
15-90C Solution
We are to discuss right-left symmetry as applied to a CFD simulation and to a potential flow solution.
Analysis In the time-averaged CFD simulation, we are not concerned about top-bottom fluctuations or periodicity. Thus, top-bottom symmetry can be assumed. However, fluid flows do not have upstream-downstream symmetry in general, even if the geometry is perfectly symmetric fore and aft. In the problem at hand for example, the flow in the channel develops downstream. Also, the flow exiting the left channel enters the circular settling chamber like a jet, separating at the sharp corner. At the opposite end, fluid leaves the settling chamber and enters the duct more like an inlet flow, without significant flow separation. We certainly cannot expect fore-aft symmetry in a flow such as this. On the other hand, potential flow of a symmetric geometry yields a symmetric flow, so it would be okay to cut our grid in half, invoking fore-aft symmetry. Discussion If unsteady or oscillatory effects were important, we should not even specify top-bottom symmetry in this kind of flow field.
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Chapter 15 Computational Fluid Dynamics 15-91C Solution
We are to discuss improvements to the given computational domain.
Analysis (a) Since Gerry is not interested in unsteady fluctuations (which may be unsymmetric), he could eliminate half of the domain. In other words, he could assume that the axis is a plane of symmetry between the top and bottom of the channel. Gerry’s grid would be cut in size by a factor of two, leading to approximately half the required CPU time, but yielding virtually identical results.
(b) The fundamental flaw is that the outflow boundary is not far enough downstream. There will likely be flow separation at the corners of the sudden contraction. With a duct that is only about three duct heights long, it is possible that there will be reverse flow at the outlet. Even if there is no reverse flow, the duct is nowhere near long enough for the flow to achieve fully developed conditions. Gerry should extend the outlet duct by many duct heights to allow the flow to develop downstream and to avoid possible reverse flow problems. Discussion The inlet appears to be perhaps too short as well. If Gerry specifies a fully developed channel flow velocity profile at the inlet, his results may be okay, but again it is better to extend the duct many duct heights beyond what Gerry has included in his computational domain.
15-92C Solution desirable.
We are to discuss a feature of modern computer systems for which nearly equal size multiblock grids are
Analysis The fastest computers are multi-processor computers. In other words, the computer system contains more than one CPU – a parallel computer. Modern parallel computers may combine 32, 64, 128, or more CPUs or nodes, all working together. In such a situation it is natural to let each node operate on one block. If all the nodes are identical (equal speed and equal RAM), the system is most efficient if the blocks are of similar size. Discussion In such a situation there must be communication between the nodes. At the interface between blocks, for example, information must pass during the CFD iteration process.
15-93C Solution We are to discuss the difference between multigridding and multiblocking, and we are to discuss how they may be used to speed up a CFD calculation. Then we are to discuss whether multigridding and multiblocking can be applied together. Analysis Multigridding has to do with the resolution of an established grid during CFD calculations. With multigridding, solutions of the equations of motion are obtained on a coarse grid first, followed by successively finer grids. This speeds up convergence because the gross features of the flow are quickly established on the coarse grid (which takes less CPU time), and then the iteration process on the finer grid requires less time. Multiblocking is something totally different. It refers to the creation of two or more separate blocks or zones, each with its own grid. The grids from all the blocks collectively create the overall grid. As discussed in the previous problem, multiblocking can have some speed advantages if using a parallel-processing computer. In addition, some CFD calculations would require too much RAM if the entire computational domain were one large block. In such cases, the grid can be split into multiple blocks, and the CFD code works on one block at a time. This requires less RAM, although information from the dormant blocks must be stored on disk or solid state memory chips, and then swapped into and out of the computer’s RAM. There is no reason why multigridding cannot be used on each block separately. Thus, multigridding and multiblocking can be used together. Discussion Although all the swapping in and out requires more CPU time and I/O time, for large grids multiblocking can sometimes mean the difference between being able to run and not being able to run at all.
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Chapter 15 Computational Fluid Dynamics 15-94C Solution We are to discuss why we should spend a lot of time developing a multiblock structured grid when we could just use an unstructured grid. Analysis There are several reasons why a structured grid is “better” than an unstructured grid, even for a case in which the CFD code can handle unstructured grids. First of all the structured grid can be made to have better resolution with fewer cells than the unstructured grid. This is important if computer memory and CPU time are of concern. Depending on the CFD code, the solution may converge more rapidly with a structured grid, and the results may be more accurate. In addition, by creating multiple blocks, we can more easily cluster cells in certain blocks and locations where high resolution is necessary, since we have much more control over the final grid with a structured grid. Discussion
As mentioned in this chapter, time spent creating a good grid is usually time well spent.
15-95 Solution
We are to calculate flow through a single-stage heat exchanger.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-96 Solution exchanger.
We are to study the effect of heating element angle of attack on heat transfer through a single-stage heat
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-97 Solution
We are to calculate flow through a single-stage heat exchanger.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
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Chapter 15 Computational Fluid Dynamics 15-98 Solution exchanger.
We are to study the effect of heating element angle of attack on heat transfer through a two-stage heat
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-99 Solution
We are to study the effect of spin on a cylinder using CFD, and in particular, analyze the lift force.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-100 Solution We are to study the effect of spin speed on a spinning cylinder using CFD, and in particular, analyze the lift force as a function of rotational speed in nondimensional variables. Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-101 Solution
We are to study flow into a slot along a wall using CFD.
Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
15-102 Solution We are to calculate laminar flow into a 2-D slot, compare with irrotational flow theory, and with results of the previous problem, and discuss the vorticity field. Assumptions
1 The flow is steady and 2-D. 2 The flow is laminar.
Analysis The flow field does not change much from the previous problem, except that a thin boundary layer shows up along the floor. The vorticity is confined to a region close to the floor – vorticity is negligibly small everywhere else, so the irrotational flow approximation is appropriate everywhere except close to the floor. Discussion The irrotational flow approximation is very useful for suction-type flows, as in air pollution control applications (hoods, etc.).
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Chapter 15 Computational Fluid Dynamics 15-103 Solution We are to model the flow of air into a vacuum cleaner using CFD, and we are to compare the results to those obtained with the potential flow approximation. V 2L
Analysis We must include a second length scale in the problem, namely the width w of the vacuum nozzle. For the CFD calculations, we set w = 2.0 mm and place the inlet plane of the vacuum nozzle at b = 2.0 cm above the floor (Fig. 1). Only half of the flow is modeled since we can impose a symmetry boundary condition along the y-axis. We use the same volumetric suction flow rate as in the example problem, i.e., V / L = 0.314 m2/s, but in the CFD analysis we specify only half of this value since we are modeling half of the flow field. Results of the CFD calculations are shown in Fig. 2. Fig. 2a shows a view of streamlines in the entire computational plane. Clearly, the streamlines far from the inlet of the nozzle appear as rays into the origin; from “far away” the flow feels the effect of the vacuum nozzle in the same way as it would feel a line sink. In Fig. 2b is shown a close-up view of these same streamlines. Qualitatively, the streamlines appear similar to those predicted by the irrotational flow approximation. In Fig 2c we plot contours of the magnitude of vorticity. Since irrotationality is defined by zero vorticity, these vorticity contours indicate where the irrotational flow approximation is valid – namely in regions where the magnitude of vorticity is negligibly small.
Far field
Symmetry line
P∞
y w Vacuum nozzle
b
x
Floor
FIGURE 1 CFD model of air being sucked into a vertical vacuum nozzle; the y-axis is a line of symmetry (not to scale –the far field is actually much further away from the nozzle than is sketched here).
V 2L
Far field
P∞ Symmetry line
Vacuum nozzle
y b
Floor (a)
y x
Floor
x
(b) 0.005
CP
V 2L
FIGURE 2 CFD calculations of flow into the nozzle of a vacuum cleaner; (a) streamlines in the entire flow domain, (b) close-up view of streamlines, (c) contours of constant magnitude of vorticity illustrating regions where the irrotational flow approximation is valid, and (d) comparison of pressure coefficient with that predicted by the irrotational flow approximation.
Irrotational flow approximation
0
-0.005 -0.01
Irrotational flow region
-0.015
Vacuum nozzle
-0.02 -0.025
b
y
Rotational flow regions
-0.03
x
Floor
CFD calculations
-0.035 -0.04 0
1
2 3 x*
4
5
(c) (d) We see from Fig. 2c that vorticity is negligibly small everywhere in the flow field except close to the floor, along the vacuum nozzle wall, near the inlet of the nozzle, and inside the nozzle duct. In these
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Chapter 15 Computational Fluid Dynamics regions, net viscous forces are not small and fluid particles rotate as they move; the irrotational flow approximation is not valid in these regions. Nevertheless, it appears that the irrotational flow approximation is valid throughout the majority of the flow field. Finally, the pressure coefficient predicted by the irrotational flow approximation is compared to that calculated by CFD in Fig. 2d. Discussion For x* greater than about 2, the agreement is excellent. However, the irrotational flow approximation is not very reliable close to the nozzle inlet. Note that the irrotational flow prediction that the minimum pressure occurs at x* ≈ 1 is verified by CFD.
15-104 Solution We are to compare CFD calculations of flow into a vacuum cleaner for the case of laminar flow versus the inviscid flow approximation. Analysis Students will have varied results, depending on the grid generation code, CFD code, turbulence model, and their choice of computational domain, etc. Discussion Instructors can add more details to the problem statement, if desired, to ensure consistency among the students’ results.
KJ
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