Handbook of Lubrication and Tribology: Volume I Application and Maintenance, Second Edition (Handbook of Lubrication (Theory & Practice of Tribology))

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HANDBOOK of LUBRICATION and TRIBOLOGY V O LU ME I Application and Maintenance S E C O N D

© 2006 by Taylor & Francis Group, LLC

E D I T I O N

HANDBOOK of LUBRICATION and TRIBOLOGY V O L U ME I Application and Maintenance S E C O N D

E D I T I O N

Edited by

George E. Totten

Boca Raton London New York

CRC is an imprint of the Taylor & Francis Group, an informa business

© 2006 by Taylor & Francis Group, LLC

Published in 2006 by CRC Press Taylor & Francis Group 6000 Broken Sound Parkway NW, Suite 300 Boca Raton, FL 33487-2742 © 2006 by Taylor & Francis Group, LLC CRC Press is an imprint of Taylor & Francis Group No claim to original U.S. Government works Printed in the United States of America on acid-free paper 10 9 8 7 6 5 4 3 2 1 International Standard Book Number-10: 0-8493-2095-X (Hardcover) International Standard Book Number-13: 978-0-8493-2095-8 (Hardcover) This book contains information obtained from authentic and highly regarded sources. Reprinted material is quoted with permission, and sources are indicated. A wide variety of references are listed. Reasonable efforts have been made to publish reliable data and information, but the author and the publisher cannot assume responsibility for the validity of all materials or for the consequences of their use. No part of this book may be reprinted, reproduced, transmitted, or utilized in any form by any electronic, mechanical, or other means, now known or hereafter invented, including photocopying, microfilming, and recording, or in any information storage or retrieval system, without written permission from the publishers. For permission to photocopy or use material electronically from this work, please access www.copyright.com (http://www.copyright.com/) or contact the Copyright Clearance Center, Inc. (CCC) 222 Rosewood Drive, Danvers, MA 01923, 978-750-8400. CCC is a not-for-profit organization that provides licenses and registration for a variety of users. For organizations that have been granted a photocopy license by the CCC, a separate system of payment has been arranged. Trademark Notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation without intent to infringe. Library of Congress Cataloging-in-Publication Data Catalog record is available from the Library of Congress

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and the CRC Press Web site at http://www.crcpress.com

This book is dedicated to my wife Alice (Ah Kum) for her continued support which allows me to pursue my passion in teaching through publication. It is her perseverance that literally defines the character of our family.

© 2006 by Taylor & Francis Group, LLC

STLE Preface

Handbook of Lubrication and Tribology: Volume I Application and Maintenance, Second Edition is sponsored and copublished by the Society of Tribologists and Lubrication Engineers (STLE). This book is one of a three volume series covering: in Volume II, Theory and Practice; in Volume III, Monitoring, Materials, Synthetic Lubricants, and Applications; and in this volume, Applications and Maintenance. The goal of this book is to provide an update to the first edition and to provide the latest information regarding application and maintenance in the broad field of Lubrication Engineering. The first edition was written over twenty years ago. In the intervening period, the science of tribology and the development of engineering best practices have evolved markedly. As a result, each of the chapters of the first edition has been rewritten and updated. A number of new chapters have been added to capture new information: In the section on Applications, the Hydraulics chapter was split into two chapters on pumps and fluids, and all new chapters on Tribology of Data Storage Devices, and Biotribology were added. In the section on Industrial Practices, a new chapter was added on Tribology of Metal Forming Processes. In the section on Maintenance, three new chapters were added on Lubricant Cleanliness, Environmental Implications of Lubricants, and Centralized Lubrication Systems — Theory and Practice. This volume was written by a peer recognized team of expert contributors from a wide variety of industry segments. Each chapter was written by an expert both knowledgeable and active in the subject area. Thanks go to these individuals; without their expertise and hard work this work would not have been possible. Thanks must also go to their employers for their support of this effort and contribution to our industry. Because of its emphasis on the practice of Lubrication Engineering, this book is an excellent reference for those preparing for STLE’s Certified Lubrication Specialist® Certification examination. As such, it has been recommended in the Body of Knowledge by STLE’s Certified Lubrication Specialist Certification Committee. This volume, like its predecessor, belongs in the reference library of all professionals in the field.

R.M. Gresham STLE Director of Professional Development

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Preface

The first edition of the Handbook of Lubrication: Theory and Practice of Tribology — Volume I: Application and Maintenance was edited by E.R. Booser and was sponsored by the Society of Tribologists and Lubrication Engineers (STLE) to provide the latest information in the field. Volume I of the Handbook of Lubrication: Theory and Practice of Tribology covers Applications and Maintenance. Volume II covers theory and design and was published in 1984. Volume III which covers Monitoring, Materials, Synthetic Lubricants, and Applications was published in 1994 to extend the topical areas covered by Volume I and Volume II since their initial publication. Over 20 years have elapsed since the First Edition was published in 1983, and enormous changes continue to occur in the lubrication and tribology engineering sciences. Although Volume III did extend the areas covered, all of the areas initially covered in Volume I needed to be significantly updated. In view of these changes and the time that has elapsed since the appearance of the First Edition, STLE initiated the Second Edition of this invaluable text. The Second Edition of the Handbook of Lubrication and Tribology: Volume I Application and Maintenance, has been reorganized slightly to aid the reader in identifying chapters and topics of interest. All of the chapters from the First Edition, with the exception of the chapter on Marine Equipment, have been revised or completely rewritten. In addition, a number of new chapters have been added including: Biotribology, Tribology of Data Storage Devices, Tribology of Metal Forming Processes, and Environmental Implications of Lubricants. The chapter on Compressors and Vacuum Pumps was significantly expanded and the original chapter on Hydraulic Systems and Fluids was divided and expanded into two separate chapters: Hydraulic Pumps and Hydraulic Fluids. Altogether there are a total of 37 chapters much of which is either a totally new treatment of the subject or completely new information. This handbook provides the reader with an extensive reference to the most important and commonly encountered lubrication systems and fluids in industry. This text is of value to the practicing tribologist and lubrication engineer, mechanical or materials engineer, and failure analysis personnel. I am indebted to all of the contributing authors of the book for their tremendous effort and patience. Without their dedication and support, the successful completion of this text would not have been possible.

George E. Totten Seattle, WA

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© 2006 by Taylor & Francis Group, LLC

Acknowledgments

I wish to thank the Society of Tribologists and Lubrication Engineers (STLE) for their continued support throughout this project. Very special thanks to Robert Gresham and Barbara Rapacz, without whose support the successful completion of this project would not have been possible. I am especially indebted to Theresa Delforn and Shelley Kronzek of CRC Press, Inc. for their continued guidance and expert assistance throughout this process, from the beginning to the end. They have made a potentially difficult task into an absolute delight. Finally, and most importantly, I am especially indebted to my family, especially my wife Ah Kum for allowing me to be so totally involved in this project for such a long time. Their continued patience with my sometimes bad behavior is most especially appreciated.

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© 2006 by Taylor & Francis Group, LLC

The Editor

George E. Totten is President of G.E. Totten & Associates, LLC in Seattle, Washington, and a visiting professor of materials science at Portland State University. Dr Totten is coeditor of a number of books including Steel Heat Treatment Handbook, Handbook of Aluminum, Handbook of Hydraulic Fluid Technology, Mechanical Tribology, and Surface Modification and Mechanisms (all titles of CRC Press), as well as the author or coauthor of over 400 technical papers, patents, and books on lubrication, hydraulics, and thermal processing. Dr Totten is a Fellow of ASM International, SAE International and the International Federation for Heat Treatment and Surface Engineering (IFHTSE) and a member of other professional organizations including ACS, ASME, and ASTM. Dr Totten formerly served as president of IFHTSE. He received bachelor’s and master’s degrees from Fairleigh Dickinson University in Teaneck, New Jersey and a Ph.D. degree from New York University, New York.

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Contributors

James R. Anglin

Dennis W. Brinkman

Sabrin Gebarin

Aluminum Company of America Alcoa Technical Center Alcoa Center, PA

Indiana Wesylan University Marion IN

Noria Corporation Tulsa, OK

José Castillo

O.-C. Göhler

Mark Barnes Noria Reliability Solutions Noria Corporation Tulsa, OK

D.J.W Barrell School of Mechanical Engineering The University of Leeds Leeds, UK

Edward P. Becker General Motors Powertrain

Bharat Bhushan Department of Mechanical Engineering Ohio State University Columbus, Ohio

Iomega Corporation Advance R&D Dept Roy, UT

Paul Conley Lincoln Industrial St. Louis, MO

J. Fisher School of Mechanical Engineering Institute of Medical and Biological Engineering University of Leeds Leeds, UK

Andy Hall Rolls Royce plc Customer Training Centre Derby, England

Hooshang Heshmat Mohawk Innovative Technology, Inc. Albany, New York

Malcolm F. Fox De Montfort University Leicester, UK

G.S. Fox-Rabinovich Department of Mechanical Engineering McMaster University Hamilton, Ontario, Canada

© 2006 by Taylor & Francis Group, LLC

Lincoln Industrial St. Louis, MO

Noria Corporation Tulsa, OK

Union Carbide Corporation Tarrytown, NY

EPRI/NMAC Consultant Lubricants of Lubrication San Rafael, CA

Ayzik Grach

James C. Fitch

Roland J. Bishop

Robert O. Bolt

Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University Aachen, Germany

Arup Gangopadhyay Ford Research Laboratory Dearborn, MI

Emile van der Heide TNO Industrial Technology Eindhoven, The Netherlands

E. Ingham School of Biochemistry and Microbiology Institute of Medical and Biological Engineering University of Leeds Leeds, UK xv

xvi

Contributors

Douglas M. Jahn

Jude Liu

Farrukh Qureshi

Delphi Saginaw Steering Systems Saginaw, Michigan

Agriculture and Biosource Engineering Department University of Saskatchewan Saskatoon, Saskatchewan, Canada

The Lubrizol Corporation Wickliffe, OH

Mark J. Jansen NASA Glenn Research Center Tribology and Surface Science Branch Cleveland, OH

Z.M. Jin School of Mechanical Engineering University of Leeds Leeds, UK

Mike Johnson Noria Field Services Noria Corporation Tulsa, OK

Robert L. Johnson Noria Corporation Tulsa, OK

Michael L. McMillan General Motors R&D Center, Chemical and Environmental Science Laboratory Warren, MI

T. Meindorf Argo-Hytos GmbH Kraichtal, Germany

Hans M. Melief The Rexroth Corporation Industrial Hydraulics Division Bethlehem, PA

Paul W. Michael William R. Jones NASA Glenn Research Center Tribology and Surface Science Branch Cleveland, OH

Milwaukee School of Engineering Fluid Power Institute, Milwaukee, Wisconsin

H. Murrenhoff Rob Dwyer-Joyce Department of Mechanical Engineering University of Sheffield Sheffield, UK

Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University Aachen, Germany

Barbara J. Parry T. Kazama Dept of Mechanical Systems Engineering Muroran Institute of Technology Hokkaido, Japan

Mohawk Lubricants North Vancouver, Canada

B.C. Pettinato Elliott Turbomachinery Co., Inc. Jeannette, PA

R. Lal Kushwaha Agriculture and Bioresource Engineering Department University of Saskatchewan Saskatoon, Saskatchewan, Canada

H.A. Poitz Air BP Lubricants Melbourne, Australia

M. Priest Roger Lewis Department of Mechanical Engineering University of Sheffield Sheffield, UK

© 2006 by Taylor & Francis Group, LLC

Jost Professor of Engineering Tribology School of Mechanical Engineering The University of Leeds Leeds, UK

Dirk Jan Schipper University of Twente Department of Mechanical Engineering Tribology Group Enschede, The Netherlands

Rick Schrama Dofasco Inc., General Maintenance Shops Hamilton, Ontario, Canada

Shirley E. Schwartz General Motors (retired)

Will Scott School of Mechanical, Manufacturing and Medical Engineering Queensland University of Technology Brisbane, Australia

Paul D. Seemuth Tribology Consulting, International LLC Hixson, TN

L.S. Shuster Department of Mechanical Engineering Ufa Aviation Institute Ufa, Russia

Jacek Stecki Subsea Engineering Research Group Department of Mechanical Engineering Monash University Melbourne, Australia

Richard K. Tessmann FES, Inc. Stillwater, OK

C.D. Tipton The Lubrizol Corporation Wickliffe, OH

Contributors

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Allison M. Toms

Drew D. Troyer

James F. Walton II

Condition Assessment Center GasTOPS Inc. Pensacola, FL

Noria Corporation Tulsa, OK

Mohawk Innovative Technology, Inc. Albany, New York

Larry A. Toms Technical Services Pensacola, FL

Simon C. Tung General Motors R&D Center Chemical and Environmental Science Laboratory Warren, MI

S.C. Veldhuis G.E. Totten Portland State University Department of Mechanical and Materials Engineering Portland, OR

© 2006 by Taylor & Francis Group, LLC

McMaster Manufacturing Research Institute (MMRI) Department of Mechanical Engineering (JHE-316) McMaster University Hamilton, Ontario, Canada

Martin Williamson Noria UK, Ltd. Chester Cheshire, UK

R.E. Yungk Air BP Lubricants Melbourne, Australia

Contents

SECTION I 1

Applications

Automotive Engine Oil

. . . . . . . . . . . . . . . . . . .

1-3

Simon C. Tung, Michael L. McMillan, Edward P. Becker, and Shirley E. Schwartz

2

Automatic Transmission Fluids . . . . . . . . . . . . . . . .

2-1

C.D. Tipton

3

Rear Axle Lubrication . . . . . . . . . . . . . . . . . . . .

3-1

Arup Gangopadhyay and Farrukh Qureshi

4

Automotive Chassis and Driveline Lubrication

. . . . . . . . .

4-1

. . . . . . . . . . . . . .

5-1

. . . . . . . . . . . . . . . . . . . .

6-1

Douglas M. Jahn and Simon C. Tung

5

Diesel, Dual-Fuel, and Gas Engines D.J.W Barrell and M. Priest

6

Aircraft Gas Turbines Andy Hall

7

Principles of Gas Turbine Bearing Lubrication and Design

. . . .

7-1

Steam Turbines . . . . . . . . . . . . . . . . . . . . . . .

8-1

Hooshang Heshmat and James F. Walton II

8

B.C. Pettinato

9

Compressors and Vacuum Pumps . . . . . . . . . . . . . . .

9-1

T. Kazama and G.E. Totten

10 Basic Hydraulic Pump and Circuit Design

. . . . . . . . . . . 10-1

Richard K. Tessmann, Hans M. Melief, and Roland J. Bishop xix

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Contents

11 Hydraulic Fluids

. . . . . . . . . . . . . . . . . . . . . . 11-1

H. Murrenhoff, O.-C. Göhler, and T. Meindorf

12 Coolants and Lubricants in Metal Cutting

. . . . . . . . . . . 12-1

S.C. Veldhuis, G.S. Fox-Rabinovich, and L.S. Shuster

13 Lubricating Industrial Electric Motors . . . . . . . . . . . . . 13-1 Drew D. Troyer

14 Effects of Radiation on Lubricants . . . . . . . . . . . . . . . 14-1 Robert O. Bolt

15 Wire Rope and Chain

. . . . . . . . . . . . . . . . . . . . 15-1

Paul Conley

16 Tribology of Hard Disk Drives — Magnetic Data Storage Technology . . . . . . . . . . . . . . . . . . . 16-1 José Castillo and Bharat Bhushan

17 Biotribology: Material Design, Lubrication, and Wear in Artificial Hip Joints . . . . . . . . . . . . . . . . . . . . . . . . . 17-1 Z.M. Jin, J. Fisher, and E. Ingham

SECTION II

Industrial Lubrication Practices

18 Steel Industry . . . . . . . . . . . . . . . . . . . . . . . . 18-3 Rick Schrama

19 Aluminum Metalworking Lubricants

. . . . . . . . . . . . . 19-1

James R. Anglin

20 Mining Industry . . . . . . . . . . . . . . . . . . . . . . . 20-1 Will Scott

21 Farm and Construction Equipment

. . . . . . . . . . . . . . 21-1

R. Lal Kushwaha and Jude Liu

22 Industrial Lubrication Practice — Wheel/Rail Tribology

. . . . . 22-1

Roger Lewis and Rob Dwyer-Joyce

23 Lubrication in the Timber and Paper Industries . . . . . . . . . 23-1 Paul W. Michael

24 Textile Fibers/Fabrics

. . . . . . . . . . . . . . . . . . . . 24-1

Paul D. Seemuth

25 Food-Grade Lubricants and the Food Processing Industry . . . . . 25-1 James C. Fitch, Sabrin Gebarin, and Martin Williamson

26 Aviation Industry . . . . . . . . . . . . . . . . . . . . . . 26-1 H.A. Poitz and R.E. Yungk

© 2006 by Taylor & Francis Group, LLC

Contents

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27 Lubrication for Space Applications . . . . . . . . . . . . . . . 27-1 William R. Jones and Mark J. Jansen

28 Friction and Wear in Lubricated Sheet Metal Forming Processes . . 28-1 E. van der Heide and Dirk Jan Schipper

SECTION III

Maintenance

29 The Degradation of Lubricants in Service Use . . . . . . . . . . 29-3 Malcolm F. Fox

30 Lubricant Properties and Test Methods . . . . . . . . . . . . . 30-1 Larry A. Toms and Allison M. Toms

31 Contamination Control and Failure Analysis . . . . . . . . . . 31-1 Jacek Stecki

32 Environmental Implications and Sustainability Concepts for Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . 32-1 Malcolm F. Fox

33 Lubrication Program Development and Scheduling

. . . . . . . 33-1

Mike Johnson

34 Lubricant Storage, Handling, and Dispensing

. . . . . . . . . . 34-1

Mark Barnes

35 Conservation of Lubricants and Energy . . . . . . . . . . . . . 35-1 Robert L. Johnson and James C. Fitch

36 Centralized Lubrication Systems — Theory and Practice . . . . . 36-1 Paul Conley and Ayzik Grach

37 Used Oil Recycling and Environmental Considerations . . . . . . 37-1 Dennis W. Brinkman and Barbara J. Parry

SECTION

Appendices

Appendix 1

. . . . . . . . . . . . . . . . . . . . . . . . . . A1-3

Appendix 2

. . . . . . . . . . . . . . . . . . . . . . . . . . A2-5

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I Applications

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1 Automotive Engine Oil 1.1

Automotive Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1-4

Engine Operation • Crankshaft to Crankshaft Bearing • Piston Pin to Piston • Piston Skirt to Cylinder Block • Piston Rings to Cylinder Block • Camshaft to Cam Follower and Valve Train • Oil Pump • Oil Filter

1.2 1.3

Issues Related to Energy Consumption in an Engine: Service Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Engine Oil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1-8 1-11

Characteristics of Engine Oil and Functions of Its Additives • Viscosity Effects • Engine Oil Quality and Oil Degradation During Vehicle Use • Fluid Film Lubrication • Future Concerns

1.4

Gasoline Engine Oil Performance Categories and Associated Test Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1-16

Introduction • ILSAC GF-4 and API SM Standard Tests

1.5

Edward P. Becker General Motors Powertrain

Shirley E. Schwartz General Motors (retired)

1-19

Service Effects of Diesel Engines Related to Engine Oil Degradation • Examples of Test Methods for Diesel Engine Oils • A Model for the Rate of Engine Oil Degradation in Diesel Engines

Simon C. Tung and Michael L. McMillan General Motors R&D Center, Chemical and Environmental Science Laboratory

Diesel Engine Oil Performance Categories and Associated Test Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1.6

Future Directions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1-21

Concerns Related to Conservation of Fuel • Effects at the Molecular Level • Insights Gained from Tests with an Alternative Fuel • Prolonging the Working Life of Engine Oil • Minimizing Emissions and Pollutants and Ensuring Backward Compatibility • Future Investigations

Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1-25 1-25

This chapter describes the functions of typical engines (gasoline and diesel), engine oil characteristics, and test methods. Included are descriptions of the tribological concerns associated with various engine components, service effects on engine oil, standard tests for engine oil and the types of service they represent, and an overview of the issues that need to be addressed in the future. 1-3

© 2006 by Taylor & Francis Group, LLC

1-4

Handbook of Lubrication and Tribology Rockers Valve springs Piston rings

Camshaft Valve Piston

Oil filter Cylinder block Con rod Journal bearings Crankshaft Oil pump Oil

FIGURE 1.1

Oil consumption Sump

The main components in an internal combustion engine.

1.1 Automotive Engines 1.1.1 Engine Operation An internal combustion engine, such as illustrated in Figure 1.1, is the predominant power source for most types of cars and trucks [1]. Various conditions influence engine development, such as the desire for high power output, legislative requirements for reduced emissions, increased fuel economy, and minimal generation of hazardous substances. Many technical and environmental challenges await those who attempt to address these concerns. The following discussion contain examples of current conditions (and in some cases the evolution of current conditions) to provide insight into the types of issues that must be understood and actions that are desirable to meet future concerns successfully. Directions in which future developments may evolve are included. Passenger car engines in North America typically use a “four stroke” cycle, which represents the number of times a piston changes direction before the events in the process of powering the engine are repeated. Some diesel and spark-ignited engines use a “two stroke” cycle, but this is not common for passenger car applications because two stroke engines may provide higher emissions of unburned fuel. Some engines locate the camshaft and valves above the engine and others locate the valves within the engine block. Engines also differ with regard to the number of cylinders and the orientation of those cylinders, such as inline or V-shaped. Figure 1.2 provides an example of a typical V-6 (six-cylinder) engine. The working mechanism of a spark ignition engine is as follows: 1. Intake: One of the valves (the intake) in the cylinder head opens when the piston is near the top of the cylinder, and as the piston moves downward, air and fuel are injected and move downward with the cylinder. 2. Compression: When the piston begins to move upward again, both valves are closed, and the contents of the cylinder (vaporized fuel and air) are compressed. 3. Power: As the piston nears the top of the compression stroke, a spark plug fires and combustion of the fuel takes place. The burning fuel creates carbon dioxide, water vapor, and other compounds. As a consequence of this gas formation, pressure rises rapidly within the cylinder. The force of the combustion gases pushes the piston down again. 4. Exhaust: As the piston reaches the bottom of the power stroke, energy from the expanding gases has been transferred from the piston to the crankshaft via the connecting rod. At this point, the exhaust valve opens and the piston then rises and sweeps most of the combustion products out of

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Automotive Engine Oil

1-5

FIGURE 1.2 Cross-section view of a V-6 engine.

the cylinder. When the piston is near the top of the exhaust stroke, the exhaust valve closes and the intake valve opens. The sequence of events then repeats. Any given engine can have various numbers of cylinders and arrangements of those cylinders, but common arrangements are V-6, V-8, and inline 4. The “cam-in-block” engine type is becoming less prevalent than an overhead cam, since a cam-in-block engine requires stiffer springs and a higher load on the springs. Lighter loads on engine components tend to reduce both wear and energy consumption. Many diverse surfaces interact and have the potential to experience wear when converting the chemical energy of the fuel into the mechanical energy of the crankshaft. These interactions are described below.

1.1.2 Crankshaft to Crankshaft Bearing As the crankshaft turns, sliding occurs between the crankshaft and the engine block structure as well as between the crankshaft and the connecting rod. The load is transferred through journal bearings, which are designed to run primarily under hydrodynamic conditions. In this bearing interface, the engine oil acts mainly as a viscous fluid, and the friction in the bearings is directly related to the viscosity of the engine oil. Since the proper clearance between the shaft and the bearing is important for good engine performance, both the shaft and the bearings should be manufactured using strong and stiff materials, to minimize deformation. Vehicle engines, however, are often shut down for long periods of time. A shaft will then settle into contact with its bearing until the engine is started again. Also, solid particles (such as residues from manufacturing, contamination, wear, etc.) can be entrained in the engine oil, and these particles have

© 2006 by Taylor & Francis Group, LLC

1-6

Handbook of Lubrication and Tribology

the potential to damage the shaft or bearing surface if the particles are larger than the minimum clearance between the shaft and bearing. Soft, compliant bearing surface materials minimize the sticking of the shaft to the bearing during shutdown, and such materials also can capture some small debris particles and remove them from circulation. This property of a bearing material to capture debris is called embedability [2]. To meet these contradictory requirements, crankshafts are usually made from a hard, stiff material such as cast iron or steel. The bearing is made using a steel backing (for strength and dimensional stability) and coated with a soft alloy (for embedability). For many years, lead-based alloys were used in crankshaft bearing applications. However, legislation now forbids the use of lead in many applications, and the low strength of the lead alloys limits the output of engines. Modern engine bearing coatings are usually made from aluminum-tin alloys, which are stronger but also have poorer embedability, so that engine and oil cleanliness become critical for long-term engine durability [2].

1.1.3 Piston Pin to Piston The piston pin transfers force from the piston to the connecting rod. The interface between the pin and the piston is also a type of journal bearing, but the motion in this case is not full rotation. In the fixed pin design, the pin is press-fit into the connecting rod, and the motion between the pin and the piston is fully reversed partial rotation. In the floating pin design, the pin is free to rotate within both the rod and the piston, and the motion is indeterminate. The floating pin has been shown to reduce the operating temperature of the piston pin boss and is therefore the preferred design [3]. In either case, the velocity of the pin is not sufficient to generate a full fluid film between the surfaces, and a condition of boundary lubrication results. The tribological properties in the pin to pin–bore interface are primarily controlled by the material properties of these parts. Automotive pistons are usually made from aluminum–silicon alloys. The piston pins are usually made from low or medium carbon steel, which is formed into a hollow cylinder and is then carburized. The carburization process results in very high hardness of the pin surface and helps minimize adhesion between the pin and piston. It has been demonstrated that increasing the oil supply to this interface reduces the tendency for scuffing [3].

1.1.4 Piston Skirt to Cylinder Block The piston skirt to cylinder block interface is one of the primary contributors to total engine friction [4]. The design challenge in this case is to maintain a small clearance between the piston and the block in order to avoid seizure, while minimizing noise and vibration [5]. The aluminum–silicon alloys used for most automotive pistons are lighter than the cast iron pistons of the past, which therefore reduces engine mass and vibration. Also, the higher thermal conductivity of aluminum helps prevent overheating of the top of the piston. Sometimes, however, the piston requires additional cooling, which is usually provided by adding devices to direct a jet of oil onto the underside of the piston. In this case, the engine oil is acting as a coolant. The cylinder bore is usually made from gray cast iron, which has a lower coefficient of thermal expansion than aluminum. This creates a design challenge, since a piston with adequate clearance at running temperature may be too loose (and hence noisy) at low temperature. To reduce friction and prevent scuffing of the piston, oil must be supplied to the cylinder bore walls. Nevertheless, the clearances are so tight that special coatings are applied to most pistons, such as nickel ceramic composites or molybdenum disulfide [6,7]. These coatings also reduce the friction in the interface of the piston rings and the piston skirt with the cylinder walls.

1.1.5 Piston Rings to Cylinder Block The piston rings function as a set of sliding seals that try to separate the combustion gases above the piston from the crankcase environment below. The most common arrangement is a set of three rings,

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Automotive Engine Oil

1-7

the upper compression ring, lower compression ring, and the oil control ring, as can be seen in Figure 1.1. The ring-block sliding interface has been estimated to account for 20% of the total engine mechanical friction [8]. Oil usually reaches the cylinder bore surface by being thrown from the crankshaft after flowing through the bearings. Some oil is necessary for the compression rings to function properly, but the oil that escapes past the compression rings is lost. The oil control ring ensures that only the necessary amount of oil reaches the compression rings. The upper compression ring experiences the highest loads and oil temperatures, and it must provide a good seal to the cylinder surface with very little engine oil. To provide acceptable durability, this ring is usually made either from nitrided stainless steel or from steel coated with molybdenum.

1.1.6 Camshaft to Cam Follower and Valve Train As the camshaft rotates, it presses against a flat or roller surface, which reciprocates to open and close the valves. The interface between the camshaft and follower is unidirectional sliding between nonconformal surfaces. Although engines are designed to provide oil to this interface, it is likely that oil will be scarce at times. For example, when starting a cold engine, the cams will begin turning before pressure is sufficient to pump oil to the top of the engine. Only a few material combinations are used successfully in this application, and even those wear sufficiently during the life of an engine to require periodic adjustment or the use of self-adjusting hydraulic elements. The severity of these various surface interactions is reduced by the presence of engine oil. The complex configuration of a typical valve train is illustrated in Figure 1.3. Since friction reduction is an important means for conserving energy and preserving nonrenewable fuel sources, techniques for reducing friction will continue to become increasingly important.

1.1.7 Oil Pump To assure an adequate distribution of lubricant through the engine and sufficient flow to maintain hydrodynamic conditions in engine bearings, automotive engines use a pressurized lubrication system. The oil flows in a circuit beginning with the sump, from which the oil is drawn into a pump. The pump then delivers pressurized oil through a filter, then to passages in the block and head, to the crankshaft and camshaft bearings, as well as to the hydraulic valve lifters in engines equipped thus. The oil is then thrown from the rotating components onto the cylinder walls, valve lifters, and other components. As the oil runs off these surfaces, gravity directs it back to the sump through passages in the head and block. Two types of pumps are commonly used, the spur gear pump and the gerotor, as shown in Figure 1.4. Examples of typical bearing and seal configurations are shown in Figure 1.5. Spur gear pumps are the older design and have the advantage of relatively quiet operation. However, the gerotor has the advantage of greater efficiency and can be made to take up less space in the engine compartment, so that most recent designs use the gerotor. The pump incorporates a relief valve for pressure regulation.

1.1.8 Oil Filter The oil filter is intended to remove potentially harmful particles from circulation. The filter element is usually either a pleated paper or metal mesh. Oil filters are rated by various tests, including industry standard methods (e.g., SAE 1858) and proprietary tests. The most commonly reported performance figures are for particle removal and flow restriction. It is desirable to have the highest level of particle removal with the lowest flow restriction. Since flow decreases as trapped material increases, filter systems generally have a bypass circuit included which opens when the pressure ahead of the filter reaches a predetermined level. This allows the oil to continue circulating if the filter becomes plugged, although contaminants are then allowed to circulate through the engine.

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Handbook of Lubrication and Tribology

(a) Valve keys Rocker arm

Valve spring retainer

Pushrod

Lifter/tappet Valve spring Camshaft

Valves

(b)

FIGURE 1.3 Valve train configuration and components: (a) valve train basic configuration; (b) valve train components — valves, seat, and guides.

1.2 Issues Related to Energy Consumption in an Engine: Service Effects The fuel provides the energy to maintain vehicle motion. However, the magnitude of the fuel consumption depends on a great number of factors. The extent to which energy is lost during operation of a given vehicle will vary with such characteristics as vehicle weight, engine type, component design, operating conditions, outside temperatures, type of terrain (flat or hilly), the number of below-freezing starts in which the engine oil never warms completely before shutdown, the viscosity of the engine oil at various

© 2006 by Taylor & Francis Group, LLC

Automotive Engine Oil

1-9 Suction

Suction

Output

Output

Spur gear

FIGURE 1.4

The two main types of pumps in internal combustion engines.

FIGURE 1.5

Engine bearings and seals.

Gerotor

operating temperatures, service history and age of the vehicle, and extent of component wear [9,10]. According to the information in Figure 1.6 (derived from an automotive database), friction in the engine, transmission, and axles represents approximately 11% of the energy consumed by a light-duty vehicle such as a gasoline-fueled passenger car. Within this 11% portion of energy usage, the piston skirt and piston rings contribute significantly to energy loss. Cooling and exhaust also represent a significant fraction of the energy loss. The severity of surface interactions between the moving components in an engine is reduced, to a greater or lesser extent, by the presence of engine oil. The oil provides different functions in different regions of the engine. Lubrication conditions are often subdivided into boundary, mixed, and hydrodynamic domains, according to the Stribeck curve (Figure 1.7), which also shows the lubrication regimes in which various engine components usually operate. Figure 1.7 indicates the relationship between the coefficient of friction (vertical axis) and a term consisting of the oil’s viscosity at a given operating temperature, multiplied by the relative difference in speed between the two surfaces, divided by the load that one surface exerts on the other. The range over which various engine components operate is indicated by the horizontal arrows. It should be noted that the vertical axis is drawn on a logarithmic scale, and the differences in friction would be greater if drawn on a linear scale. The low point on Figure 1.7 indicates the condition under which friction is a minimum

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology

Distribution of energy losses in a typical light-duty vehicle

Exhaust 33%

Wheels 12%

Piston skirt friction 25%

Axle & transmission 22.5%

Air pumping 6% Braking & coasting 7.5% Engine friction 7.5% Axle & transmission friction 3% Accessories 4%

Crankshaft 5% Piston rings 19%

Valvetrain 6%

Cooling 29%

Bearings 22.5%

Research and development center

FIGURE 1.6

Typical values for energy loss in a light-duty vehicle. Boundary Coefficient of friction

1.

Mixed

Hydrodynamic

Piston rings

0.1

Piston skirt Valve train .01

Engine bearings

.001 Viscosity × Speed Unit load

FIGURE 1.7

Stribeck diagram, including the operating regions of several engine components.

(and thus fuel consumption will be minimized for a given vehicle). Engines do not operate at a constant temperature, vehicles sometimes drive on rough roads, and various additional conditions influence vehicle operation, so that Figure 1.7 represents a highly idealized assessment of friction effects. In the hydrodynamic region, the sliding surfaces are completely separated by an oil film, and friction is essentially due to shearing of the fluid. As the sliding speed and viscosity of the engine oil decrease and loads increase, the two opposing solid surfaces begin to interact. Moving to the left on the Stribeck curve, the coefficient of friction rises sharply as the load is shared between the fluid and the solid surfaces in the region identified as mixed lubrication. At some sufficiently low value of viscosity and component speed and at a sufficiently high load, the contact zone moves into the domain of boundary lubrication [1,2].

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Automotive Engine Oil

1-11

1.3 Engine Oil 1.3.1 Characteristics of Engine Oil and Functions of Its Additives The major component of engine oil is its base stock (i.e., the oil itself). Various additives are added to the engine oil, each of which provides a highly specific mode of action to protect the engine and reduce the rate at which the engine oil degrades. Zinc dialkyl-dithiophosphate (known as ZDP) is an essential additive in engine oil, and it has two functions: inhibition of oil oxidation and protection against wear. To protect the oil against oxidation, ZDP tends to react faster with oxygen than the rate of attack by oxygen on the oil base stock. In this way, the oil base stock and its other additives are less likely to be oxidized. In addition, ZDP reacts with iron on an engine’s surface (particularly in a heavily loaded contact) by laying down a phosphorus and sulfur coating that is resistant to wear. The phosphorous in ZDP can poison catalytic converters, which has contributed to a trend in recent years to reduce ZDP concentrations in engine oil. Thus, additional compounds that provide supplemental oxidation protection are generally also incorporated into an engine oil formulation. A detergent in the engine oil behaves somewhat like a soap, in that it reduces the tendency of partially oxidized oil to form tar-like deposits on a hot surface. A dispersant helps keep degraded oil from coagulating, so that the coagulated oil will not be able to block narrow lubricant passageways. A pour-point depressant allows the oil to flow at low temperature.

1.3.2 Viscosity Effects Appropriate engine oil viscosity is essential for satisfactory engine performance, but maintaining suitable viscosity over a temperature range that can extend well below 0◦ C and well above 100◦ C requires an additive in the engine oil (a viscosity index improver, typically called a “VI” improver) that helps to minimize the adverse consequences of large temperature fluctuations. A VI improver is a long-chain polymer that is less soluble in cold oil but more soluble in warm oil. When cold, the VI improver folds in upon itself and offers less resistance to oil flow. Thus, the VI improver facilitates cold starting of an engine. When the oil is hot, the VI improver expands into a loose coil, so that the viscosity of the engine oil increases over what it would otherwise be at an elevated temperature. This expansion and contraction effect may diminish as the VI improver ages and is broken down by high-shear conditions, which are likely to be experienced whenever engine oil passes through narrow, hot contact points such as in a heavily loaded bearing or underneath the piston rings. The oil container will display a term such as SAE 5W-30, in which the “5W” signifies the oil viscosity when the oil is cold; the “30” indicates the viscosity at normal operating temperatures. Viscosity requirements under various shear conditions for the different viscosity grades are established by the Society of Automotive Engineers (SAE) and are included in SAE J300, “Engine Oil Viscosity Classification.” The latest requirements are summarized in Table 1.1. Oil containers usually also display something relating to the performance capabilities of the oil. The two most common symbols indicating that engine oil satisfies a particular performance standard for gasoline engines in the United States are the API Certification Mark (starburst) and the API Service Symbol (donut).

1.3.3 Engine Oil Quality and Oil Degradation During Vehicle Use A container of engine oil, such as one would buy in a store or at a service station, should have a symbol that indicates whether the oil meets current standards (as indicated above). Unfortunately, some stores also carry engine oils that do not have a current designation, and an uninformed purchaser is at risk of buying an inappropriate grade of engine oil. In addition, overly degraded engine oil puts an engine at risk of damage. Examples of oil analysis tests that are helpful in determining the extent of engine oil degradation during use include changes in viscosity as described in various standard tests such as ASTM D 445, D 446, D 4683, and D 4684. (Note: all American Society for Testing and Materials [ASTM] standards cited in this

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1-12 TABLE 1.1

Handbook of Lubrication and Tribology SAE Viscosity Grades for Engine Oilsa

Low-temperature ◦ C cranking viscosityb (mPa sec, max)

Low-temperature ◦ C pumping viscosityc (mPa sec, max with no yield stress)

Low-shear-rate kinematic viscosityd (−mm2 /sec at 100◦ C, min)

Low-shear-rate kinematic viscosityd (mm2 /sec at 100◦ C, max)

High-shear-rate kinematic viscositye (mPa sec at 150◦ C, min)

6,200 at −35 6,600 at −30 7,000 at −25 7,000 at −20 9,500 at −15 13,000 at −10 — — —

60,000 at −40 60,000 at −35 60,000 at −30 60,000 at −25 60,000 at −20 60,000 at −15 — — —

3.8 3.8 4.1 5.6 5.6 9.3 5.6 9.3 12.5

— — — — — — 5.0 cSt at 100◦ C used from Cycling Test

DEXRON®-III

DEXRON®-IIIH

Ford Type F

None specified

None specified

fluid fluid fluid fluid fluid fluid

M2C138-CJ M2C166-H MERCON®

49 SUS min. (7.0 cSt) at 210◦ F 7.0 cSt min. at 98.9◦ C 6.8 cSt min. at 100◦ C 6.8 cSt min. at 100◦ C

46.5 SUS (6.2 cSt) min. after 8000 cycle WOT test 6.2 cSt min. after FTLM BJ 12-4 6.0 cSt min. after FTLM BJ 12-4 5.0 cSt min. in GM Cycling Test

MERCON® V

6.8 cSt min. at 100◦ C

6.0 cSt min. at 100◦ C after 20 h KRL Shear

DaimlerChrysler (Chrysler Group) MS-3256

49 SUS min. at 210◦ F

48 SUS min. at 210◦ F after Chrysler shear 461C-112

MS-4228

49 SUS min. at 210◦ F

48 SUS min. at 210◦ F after Chrysler shear 461C-112

MS-7176, Change G (ATF+3®) MS9602, Change F (ATF +4®)

7.4–7.7 cSt min. at 100◦ C 7.3 to 7.8 cSt at 100◦ C

6.5 cSt min. after 30 passes in ASTM D3945B 6.5 cSt min. after 20 h KRL Shear

7000 SUS max. (extrapolated from 210◦ F and 100◦ F values) 4500 cP max. at −10◦ F 64,000 cP max. at −40◦ F 4000 cP max. at −10◦ F 55,000 cP max. at −40◦ F in Brookfield viscometer 4000 cP max. at −10◦ F (−23.3◦ C) 50,000 cP max. at −40◦ F in Brookfield viscometer 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 5000 cP max. at −30◦ C 20,000 cP max. at −40◦ C 1400 cP max. at −40◦ F in Ford test method BJ 3-2 1700 cP max. at −18◦ C 1700 cP max. at −18◦ C 1500 cP max. at −20◦ C 20,000 cP max. at −40◦ C 1500 cP max. at −20◦ C 9000 ± 4000 cP max. at −40◦ C 7000 cP max. at −20◦ F in Chrysler method 461C114 2300 cP max. at −20◦ F in Chrysler method 461C114 4500 cP max. at −28.9◦ C 20,000 cP max. at −40◦ C 3000 cP max. at −28.9◦ C 10,000 max. at −40◦ C

a Viscosity of ATF was originally specified in Saybolt Universal Seconds (SUS) at Fahrenheit temperatures until the specific-

ations underwent metrification and were modernized. b Generally, low temperature viscosity is by the ASTM D2983 Brookfield procedure unless otherwise stated; temperatures

are stated in either Fahrenheit or Celsius depending on the time period of the specification.

© 2006 by Taylor & Francis Group, LLC

Automatic Transmission Fluids TABLE 2.2

Fluid ATF #1

ATF #2

2-5

Selected Physical Properties of Automatic Transmission Fluids

Temperature ◦ F/(◦ C)

Thermal conductivity Cal/sec cm2 (◦ C/cm) × 105

Volumetric thermal expansion ◦ C−1 × 103

Heat capacity Cal/g◦ C

Specific gravity

0/(−17.8) 100/(37.8) 200/(93.3) 300/(148.9) 0/(−17.8) 100/(37.8) 200/(93.3) 300/(148.9)

30.6a 30.4 30.2 30.0a 31.4a 31.2 31.0 30.8a

8.8 25.3 41.6 58.2 9.9 25.6 41.0 56.7

0.46a 0.50 0.55 0.60 0.43a 0.48 0.53 0.58

0.8655b 0.8584 0.8335 0.7986 0.8640b 0.8574 0.8314 0.7969

a Extrapolated values. b Extrapolated from 4◦ C.

of bands. Surface finish and hardness are critical factors of steel reaction members in maintaining proper friction and durability. A shifting clutch may experience very high temperatures, up to 600◦ C, in some cases, therefore flow of fluid through the clutch to provide cooling is very important. Often times, grooves are cut or embossed into the friction material to enhance lubrication and cooling. Composite plates are manufactured to provide compressibility and conformability to increase friction coefficient; porosity aids in both cooling and lubrication of the interface by allowing absorbed fluid to be squeezed out during engagements. With high interfacial temperatures being generated, it is important that the ATF have good thermal and oxidative stability and be able to retard buildup of decomposition products which may clog pore structures critical to the function of the friction material and “glaze” the surfaces. In a clutch engagement, the observed friction is a combination of hydrodynamic and asperity friction [10]. In the initial stage of engagement, there is mainly hydrodynamic friction brought about by shear of the fluid film. Later in an engagement one finds that asperity friction begins to take over. The automatic transmission fluid can mediate the friction coefficient observed in both phases of the engagement. Particularly, the coefficient of friction can be controlled in the very slow speed end of an engagement, where asperity contact is dominant, by the use of friction modifiers. The friction vs. speed curve (µ–v) can be controlled to having either a negative or a positive slope in this region. This is important because perceived shift quality or “shift-feel” and also the propensity for stick-slip or “shudder” to occur depends, to a large extent, on the slope of the µ–v curve of the ATF. Negative slope behavior can lead to shudder in torque converter clutches and harsh shifts in shifting clutches. The dynamic torque provided by an ATF is important to the determination of the torque capacity of the clutch unit and to the number and types of friction elements used. Static torque or friction coefficient is also important in that the static holding capacity of a clutch is dependent on this value and it is generally desired that the value be as high as practical in concert with good antishudder characteristics. Current North American ATF specifications such as DEXRON®-III (Revision H), MERCON® V, and MERCON® are for service-fill of General Motors and Ford vehicles, respectively. Factory-fill requirements generally reflect service-fill requirements but often with specific additional testing required. Friction tests used to determine the properties of an ATF are usually conducted in an SAE#2 electric motor driven friction tester using a clutch pack or band, or in a vehicle when antishudder and shift quality are of concern. SAE#2 Tests are generally of the engagement type where the clutch and fluid are cycled for up to 30,000 cycles with key parameters like dynamic torque are required to stay within narrow limits. Figure 2.2 is a diagram of a typical ATF SAE#2 engagement torque trace and the parameters used by various passenger car manufacturers to estimate performance. Slow speed or “static” friction measurements (Figure 2.3) are usually determined using an auxiliary slow speed motor to turn the plates at a constant slow speed, typically 0.72 rpm, while engagement pressure is maintained at a constant value.

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Handbook of Lubrication and Tribology

4000

300

3500

250

Apply pressure

200

3000 2500

Torque

2000

150

1500

100

1000

Speed

50

500 0 1500

0 0

Speed (rpm)

Torque (Nm) Temperature (°C) pressure (kPa)

Dynamic friction engagement

500

1000 Time (msec)

FIGURE 2.2

Dynamic friction engagement.

Static friction break-away 0.72 rpm motor speed 200

0.180

180

0.160

160

Torque (Nm)

120

Tt (mt) Torque 2 sec after the start of static trace (2175 msec)

100

0.120 0.100 0.080

80 0.060

60

Friction coefficient (Mu)

0.140 Ts (ms) Maximum torque immediately after start of static trace

140

0.040

40 5 Nm threshold start of static trace (175 msec)

20

0.020

0 0

500

1000

1500

2000

0.000 2500

Time (msec)

FIGURE 2.3

Static friction break-away.

2.4 Oxidation Stability The factory-fill automatic transmission fluid is now expected by automobile manufacturers to last for the useful life of the vehicle, a so-called fill-for-life fluid. The ATF and its frictional and viscosity characteristics are a key part of maintaining the designed shift quality and antishudder durability of the vehicle. Therefore, it is desired that the ATF not be changed such that there is little chance that a misapplication of the wrong type of ATF can occur. Drain plugs and dipsticks are being eliminated to reduce interference with the factory-fill fluid. In order to ensure that a fluid can last the life of the vehicle, good stability toward oxidation is essential. Oxidation of the fluid can (1) increase both the high temperature and the low temperature viscosity of the fluid; (2) cause the formation of sludge, varnish, and particulate matter which can interfere with the function of the friction materials and the other mechanical surfaces; (3) cause corrosion of bushings, thrust washers, and bearings; and (4) cause degradation of elastomeric seals. Oxidation of the fluid is also a function of the conditions that the fluid experiences during service in a vehicle with temperature being

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Automatic Transmission Fluids

2-7

the key factor. Oxidation in vehicles in the field has been related to the length of service time the fluid has been at high temperature. These conditions typically occur during trailer towing or in traffic at engine idle speeds. Requirements for oxidation resistance in transmission fluid specifications have become ever more stringent as equipment makers strive to ensure fill-for-life. For example, General Motors uses a full-scale electric motor-driven transmission for its oxidation testing in its DEXRON® series of specifications. From the original DEXRON® to the current DEXRON-III (H revision), the length of the oxidation test has increased from 300 to 450 h at 163◦ C and the air flow rate has been changed from 30 to 90 cc/min. Additional requirements have been added besides sludge and varnish limitations including a 3.25 maximum change in total acid number, and an infrared absorbance maximum increase of 0.45. Other tests used to measure oxidation are the Aluminum Beaker Oxidation Test (ABOT) used by Ford in its MERCON® series of specifications, ISOT (Indiana Stirring Oxidation Test) used primarily in Japan, and the DKA oxidation test (CEC L-48 METHOD) used in Europe.

2.5 Foam Excessive foam buildup in a transmission can be very detrimental to transmission function and to safety. Foaming can be induced by the churning of the rotating parts and gears in the fluid, air leakage on the suction side of the hydraulic pump, and other factors. Since one of the functions of an ATF is as a hydraulic fluid, foaming causes a large increase in the compressibility to the ATF, interfering with the pressure applied to clutches. In some cases this can lead to premature failure of the clutch due to lack of proper apply pressure. Foaming also causes the heat capacity of the ATF to be reduced, so it is less effective in removing heat from the clutches and other parts, and foaming also leads to an increase in the volume of the ATF. This latter effect can be very dangerous in vehicles having a dipstick, because if the volume of the ATF exceeds the free air space in the transmission, the ATF can be thrust up the dipstick tube and ejected.

2.6 Compatibility with Organic Materials Seals, engineering polymers, and wire insulation for electronic controls are some of the materials that are of concern when in contact with ATF. Seals are used as gaskets, shaft seals, and seals for the clutch pistons and drums that are hydraulically actuated. Deterioration of these elastomeric seals leads to either internal or external leakage in the transmission. Internal leakage can lead to the loss of adequate hydraulic pressure to critical clutches, and external leakage is a negative to the vehicle purchaser. A wide variety of different types of elastomers are used in transmission including silicone, nitrile, polyacrylate, fluorocarbon types (e.g., Viton®), and others. ATF can interact with a seal material in several ways which can be detrimental. It can either swell or it can shrink the elastomer. Too much shrinkage or too much swelling can lead to deterioration of dimensional and physical properties of the seal and cause leakage. For any ATF formulation that gives an inadequate amount of swelling (a particular problem with highly refined Group III and Group IV base fluids), chemical seal swell agents can be added to increase the positive volume change of elastomers, especially with nitrile-type elastomers. Chemical additives in the ATF can also react with elastomers, depending on the elastomer-type. For example, too much sulfur in an ATF can cause hardening of a nitrile rubber due to additional cross-linking of the polymer. Silicone rubber can be depolymerized by strong acids generated from oxidation of the fluid, especially at higher temperatures. Transmission fluids are generally tested for seal compatibility by soak tests such as the ASTM D471. A small amount of positive volume change is usually desirable in these tests, but excessive volume increase and shrinkage are to be avoided. In addition to the bench scale soak tests, parts from full-scale transmission tests are examined to determine whether each type of seal material is performing as well as they do with a good reference fluid.

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Handbook of Lubrication and Tribology

Other organic polymers are used for insulation of the electrical wires for electronic controls. These are usually tested by the wire suppliers against reference ATFs to ensure compatibility. Polymeric engineering materials found in transmissions include parts made from phenolic resins, polyesters, and nylon. Nylon has been used to fabricate gears and thrust washers used in automatic transmissions. Degradation of nylon parts can be a problem when in contact with ATF at higher temperatures [11] and either the ATF or the nylon formulation is improperly selected.

2.7 Corrosion Protection Automatic transmissions contain parts made from a variety of metals. The cases and valve bodies may be aluminum; shafts, gears, and clutch drums may be steel alloys; thrust washers and bushings can be made of various copper containing alloys; oil coolers may be made of brass; and bearings and thrust washers may be coated with tin or lead. In some heavy-duty transmissions, the clutch plates themselves may be made of sintered bronze. It is critical to the function of these types of parts that they not undergo deterioration through chemical attack from water, oxygen in the air, or from the ATF itself. Most ATFs contain specific anticorrosion agents to retard the rusting of ferrous alloys, and the corrosion of copper, zinc, lead, and tin. Among the tests used to measure anticorrosion for ATF are the ASTM D665 and the ASTM D1748 humidity cabinet for rust of ferrous alloys, and the ASTM D-130 for copper corrosion. In addition, full-scale transmission test parts are examined for corrosion and compared to good performing reference fluids. Ford uses copper and lead strips in its ABOT test and has a requirement on the maximum percentage weight loss of the lead strip.

2.8 Antiwear Automatic transmission fluid must be able to successfully lubricate and prevent wear in multiple mechanical parts in a transmission including vane and gear pumps, one-way clutches, planetary gearsets, bushings, bearings, thrust washers, and power transfer chains. Modern automatic transmissions have undergone an evolution in weight reduction and increased power density in an effort to save weight and reduce size for better vehicle fuel economy. These changes and also the desire for the transmission to last the lifetime of the vehicle with its original fluid have raised additional concern about the antiwear capability of the transmission fluid. Earlier specifications may have included only one specific wear test such as a 4-ball wear test [12,13] or a power steering vane pump wear test [4]. It is now typical for an ATF specification to include several wear tests, including a FZG four square gear test [14] and a test to measure wear of a sprag-type one way clutch. Transmissions run in dynamometer tests are also monitored for any signs of unusual wear when tested with candidate ATF fluids. Many earlier ATFs relied on zinc dithiophosphate antiwear technology, similar to that used in engine oils, for wear protection but few, if any, ATFs formulated after about 1980 in North America contain zinc dithiophosphate. ATFs nearly universally use metal-free antiwear agents employing a combination of phosphorus, boron, and sulfur materials.

2.9 Composition Automatic transmission fluids are composed of a base fluid plus a complex additive formulation intended to meet all of the physical and performance properties required of an ATF. The base fluid is usually a petroleum based or synthetic hydrocarbon mixture with a viscosity of between 3.0 and 4.5 cSt at 100◦ C. Viscosity at low temperature, volatility, and oxidation stability are important criteria in the selection of base fluid. The additive is required to help the ATF meet the physical and performance properties that are not easily met by the base fluid alone. The classes of additive components that can be used to make an ATF include

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Automatic Transmission Fluids

2-9

viscosity modifiers, fluidity modifiers, pour point depressants, foam inhibitors, dispersants, detergents, antiwear agents, oxidation inhibitors, corrosion inhibitors, detergents, and friction modifiers. Often there is more than one representative of each class present, so the total number of additive components used to make an ATF is usually in the 10 to 20 range. Viscosity of an automatic transmission fluid is a combination of the contributions from the base fluid plus the additives. Viscosity modifiers are used to adjust the high temperature viscosity and also to control and reduce the low temperature viscosity. Typically, polymethacrylate and esterified maleic anhydride/styrene copolymers are used. These high molecular weight molecules are constructed using special alcohol mixtures in the ester functions to provide modification of wax structures that may come from the base fluid. This provides improved flowability at low temperatures as measured by the Brookfield viscosity test. The viscosity modifiers also typically incorporate dispersant functionality to aid in sludge dispersancy during oxidizing conditions. Low temperature viscosity is often further controlled by the addition of special low viscosity oils and by the use of pour depressants. These pour depressants are similar to the viscosity modifiers, but are specially designed to modify the formation of wax networks and the treatment levels are lower. Specialized dispersants, usually of the polyisobutenylsuccinimide-type, are another major component of an ATF additive system. Dispersants are truly multifunctional components which act to suspend dirt, sludge, and debris, they extend clutch frictional durability through their cleaning effect on composite friction materials, and they influence friction coefficients in relation to sliding speeds. Some dispersants can also act as carriers for other additive functionalities like boron, which can be active on antiwear and anticorrosion. Metallic detergents like calcium and barium overbased sulfonates and carboxylates were once extensively used in automatic transmission fluids to provide stable friction coefficients and to neutralize acidic products of oxidation. These detergents act in much the same way as a dispersant on composite friction materials to maintain cleanliness and retard buildup of deposits and glaze. Detergents also, through their ability to neutralize acidic materials, are able to protect elastomeric seal materials and composite friction materials from premature deterioriation [15,16]. The classic antiwear agents used in automatic transmission fluids were zinc dithiophosphates. However, in the mid-1980s with the increase in the severity of clutch friction testing, the thermal instability of most zinc dithiophosphates led to the discontinuation of use in most ATF formulations. Instead, phosphorus materials which had relatively low activation temperatures, such as trivalent phosphorus esters and phosphorous acid itself, came into use for antiwear. Active sulfur compounds are generally not used in ATF formulations because of corrosiveness toward copper-based alloys, attack of elastomers used for sealing, and deleterious interactions with composite friction materials. Boron often finds use in ATF formulations as a noncorrosive antiwear supplement to the phosphorus materials used in ATF. Oxidation inhibitors are important in maintaining the properties of the ATF for the life of the vehicle. Exposure to the oxygen in the air and long periods of time at elevated operating temperatures lead to attack on the molecules of the base fluid and additives by oxygen and subsequent decomposition to oxidation propagating peroxy radicals. Oxidation inhibitors act to intercept these radicals and render them less harmful so they do not continue to propagate creating acidity and polymeric degradation products [17]. Alkylated diaryl amines is the most ubiquitous class of oxidation inhibitors found in ATFs, but certain less active sulfur compounds and hindered phenols are also used. Again, these inhibitors are used either singly or in combinations to get the most effective combination for a given additive and base fluid combination. Great care must be exercised in the selection of oxidation inhibitors in order that a low level of undesirable corrosiveness, interaction with composite friction materials, and elastomeric seals is achieved. One of the key components in an ATF formulation is the friction modifying chemistry needed to achieve the proper levels of static and dynamic friction, µ/v slope, and long term antishudder durability [18]. In today’s vehicles, there is a need to have both a relatively high static coefficient of friction for clutch holding capacity and also a positive µ/v slope for the life of the fluid. Therefore, many older classic friction modifiers which were very effective at depressing static coefficients of friction are no longer as useful. In more modern fluids, thermal and oxidative stability of the friction modifier is an important

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Handbook of Lubrication and Tribology

selection criterion. In addition, friction modifiers are chosen which produce a smaller reduction in static coefficient values. The friction modifiers normally have a polar functionality in the molecule which interacts by surface adsorption on the composite friction material and the reaction member. The oleophilic remainder of the molecule extends into the fluid from the surface [19]. It is these surface films and their structure and lability that changes the friction when the oil film becomes thin and asperity contact of the rubbing elements becomes significant. Typical functional groups that interact with the surfaces are alcohols, amines, amides, carboxylic acids, imidazolines, and oxazolines. Typical carbon chain lengths for friction modifiers are from 10 to 24 carbon atoms. Corrosion inhibition and general compatibility with materials is required for the long-term durability and function of transmission components. For ferrous metals calcium overbased sulfonates, nonionic surfactants, and certain aminic surface-active agents are employed, if necessary to pass antirust requirements. Copper, tin, zinc, and lead alloys are often used for bushings, bearings, thrust washers, electronics, coolers, cooler lines, and braze joints. Typical inhibitors of corrosion of these materials are various commercially available benzotriazole derivatives and derivatives of 3,5-dimercaptothiadiazole. Calcium overbased sulfonates act to protect metals, elastomers, and engineering plastics from acidic attack. With the move to highly refined mineral oil and synthetic hydrocarbons for base fluids, there are fewer aromatic and naphthenic components in the base fluids to maintain the volume of the elastomeric seals. In recent years, increasing amounts of seal swelling additives have been required to maintain positive swell values required by transmission manufacturers in seal immersion testing. In the past, highly aromatic oils were often used as seal agents, but the use of these ceased when carcinogenicity questions arose in the use of these materials. Today, chemical seal swell agents are often used. These are materials that have an affinity for the elastomeric seal material and can be preferentially absorbed into the interstices of the elastomer matrix. Nitrile rubbers are often the most difficult to swell, so seal swell agent treatment levels are optimized to the requirements of this type of elastomer. Typically, pthtalate esters, sulfones, or other low molecular weight esters are used in ATFs as seal swell agents at the minimal levels required to provide the specified volume change. If a naphthenic oil with good oxidation stability and other properties is available, it may also be incorporated as a seal swell agent. Other types of seal materials that are used in transmissions include fluorocarbon, polyacrylate, and silicone based elastomers, but these types of elastomers generally have less need for seal swell agents than do nitrile elastomer-types. Antifoam additives are required to prevent excessive foaming of the fluid in the transmission. Siliconetypes are most commonly used including polydimethylsiloxanes, fluorosilicones, and functionalized siloxanes. Most of these liquid materials are insoluble in oil and are incorporated into automatic transmission fluids as very fine dispersions. The silicone droplets are surface active and act to destabilize the air/oil interface of a foam bubble through changing the surface tension. Most transmission fluids are required to be dyed red by manufacturer specifications in order to identify sources of leakage and, generally, to provide fluid identification and to distinguish ATF from undyed engine oils.

2.10 Specifications and Testing Requirements The first specification to standardize the requirements for a fluid for automatic transmissions was introduced in 1949 by General Motors and was designated “Type A.” The Type A specification also provided a process for qualification of service-fill ATF through a trademark and licensing procedure, a process that continues to the present day and serves to help police and enforce the quality of transmission fluids bearing the trademarks. Eventually it was found that some Type A qualified automatic transmission fluids showed deficiencies in oxidation resistance and this led to the introduction of the Type A, Suffix A specification in 1957. In 1967, General Motors introduced the DEXRON® trademark and specification to continue to upgrade the fluids on friction durability, oxidation stability, and viscosity properties. The DEXRON® specifications have continued to be upgraded in the ensuing years.

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Automatic Transmission Fluids

2-11

In the late 1950s, Ford’s factory-fill fluids met specification M2C33-A-B which describe a fluid with similar characteristics to Type A, Suffix A. Ford introduced the M2C33-D specification in 1961. This change was driven by the need for better oxidation control, antiwear performance, and higher static capacity. Ford introduced a new type of specification for automatic transmission fluid in 1967. Ford had the objective of a “fill-for-life” fluid with improved antioxidation, wear, and friction performance. This new fluid was described by the M2C33-F specification and was conventionally called “Type F” [5]. It was also a service-fill specification and Ford granted approvals under the specification along with qualification numbers. This specification was similar to M2C-33D. Differences included a 6-pack clutch friction test which required a high static coefficient of friction. Ford’s use of an ATF with a friction curve similar to base oil would undergo less frictional change with time than a fluid that used a chemical friction modifier which might degrade, allowing the oil revert to base oil characteristics. Another driving force was to reduce the number of plates in the clutch pack to get a more consistent shift characteristic. Oxidation resistance was increased by raising the temperature of their 300-h transmission oxidation test from 300 to 325◦ F. Ford moved away from the Type F concept of a high static friction fluid with the introduction of the M2C138-CJ specification in 1974 [20], which called for a friction-modified fluid. This type of fluid was intended to alleviate difficulties with engineering for good gear engagement noise perception with the high static friction fluid. A new friction test was introduced which called for a narrow band of acceptability (0.9 to 1.0) in the ratio of static to dynamic friction in the M2C138-CJ fluids. This required that the static friction had to be less than the dynamic friction. This specification was followed by the M2C166-H [21] specification for factory-fill fluids requiring improved friction characteristics for lock-up torque converters for factory-fill fluids. In this specification, Ford introduced the ABOT to replace the 300 h oxidation test carried out in a motored transmission. The MERCON® specification was introduced by Ford in January 1987 [22], a trademarked fluid, for service-fill with procedures for qualification and licensing of fluids to ensure quality in the marketplace. The development of the modulated and continuously slipping clutch torque converters prompted the need to develop the MERCON®V specification. Requirements for verifying antiwear capabilities and antishudder characteristics were included in the specification. In the late 1960s Chrysler Corporation also upgraded its requirements for automatic transmission fluid with the MS-4228 factory-fill specification [6] replacing the older MS-3256 which had described fluid similar to Type A, Suffix A. Very good low temperature viscosity was a major objective in order to reduce the time for engine starting and the time for shifts to occur when the fluid was still cold. Requirements for oxidation stability were also upgraded through the use of a new bench oxidation test. The next major upgrade by Chrysler Corporation, now DaimlerChrysler, was the introduction of the MS-9602 specification in the late 1990s and the marketing of the new factory-fill fluid as ATF+4® [23]. A history of these specifications is given in Table 2.3.

2.11 Timeline of ATF Specifications ATF physical properties and tests were covered earlier in this chapter. Performance testing of ATFs to meet factory-fill and service-fill qualification requirements are handled by each individual OEM in the case of ATF. There is no overall specification for an ATF which can be used in any vehicle as exists for engine oils. Friction and friction durability, oxidation resistance, and antiwear characteristics are the three key areas at which most performance testing is directed. In the absence of an industry-standard specification, certain OEMs, for example, General Motors and Ford, have written their own service-fill specification and assigned a trademark designation to it, DEXRON® in the case of General Motors and MERCON® in the case of Ford Motor Company. It is through the use and licensing of the trademarks that the composition and quality of approved fluids is maintained. In each case specific tests are required that ensure suitability for service of current models and back application to previous transmission. Since not all transmission manufacturers use the same transmission designs or the same friction materials, the specifications

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Handbook of Lubrication and Tribology TABLE 2.3

Timeline of ATF Specificationsa

Specification Type A Type A, Suffix A Type F MS-3256 MS-4228 DEXRON® DEXRON®-IIb M2C138-CJ DEXRON®-II MS-7176, Change D (ATF +2®) M2C166-H MERCON® DEXRON®-IIE DEXRON®-IIE-Rev. DEXRON®-III MERCON® V DEXRON®-IIIGc MS-7176, Change G (ATF +3®) DEXRON®-IIIH MS-9602, Change F (ATF +4®)

Number

6137-M 6137-M

6137- M 6137- M 6297- M 6417- M

Year 1949 1957 1959 1964 1966 1967 1973 1974 1978, July 1980, May 1981, June 1987, January 1990, October 1992, August 1993, April 1996, July 1997, April 2001, June 2003, April 2003, July

Company GM GM Ford Chrysler Chrysler GM GM Ford GM Chrysler Ford Ford GM GM GM Ford GM DaimlerChrysler GM DaimlerChrysler

a Not all specification changes are included. b DEXRON®-II was originally released using “C” qualification numbers. In 1975 after suppliers

rolled over approvals because of a GM mandated fix for a cooler corrosion problem, the qualification numbers were preceded by a “D.” This led to referring to the fluids as DEXRON®II “C” or “D.” c Upgrade for ECCC vehicle test and sprag clutch wear.

may differ in test requirements. However, current automatic transmission fluid requirements are similar enough that it is common to have fluids approved for both a DEXRON®-type and a MERCON®-type requirement. This means that the fluid so designated has been through two separate qualification programs in order to bear both trademarks. Factory-fill fluids are usually similar to service-fill fluids in performance, but they have gone through additional testing at the manufacturer that has not been required for most service-fill fluids. Most manufacturers make their factory-fill fluid available for service use through their dealerships. Specifications for on-highway automatic transmissions generally call for friction-modified (low static coefficient of friction) fluids today. Ford-Type F fluid used prior to about 1980 was an older fluid that did not contain friction modifier. Fluids designed to meet this specification are still commercially available but are not suitable for modern Ford vehicles. Each transmission or vehicle manufacturer has a special set of requirements designed around the characteristics of the materials used in the manufacture of the transmission. Friction testing usually can be categorized into two types, initial friction characteristics tests and tests measuring durability characteristics. Both General Motors and Ford have long-term durability tests run in a SAE#2 clutch friction tester in their specifications for service-fill. These tests use flat disk-type clutch plates. The exact length of the test, the friction material used on the friction plate, the arrangement of the plates in the clutch pack, temperatures, and so on are unique to the specification involved. Also, the acceptable limits in the envelope of change of the test parameters allowed throughout the test vary. Table 2.4 shows a comparison of the test parameters for three major service-fill ATF specifications. Other long-term friction tests include a band friction durability test and the engine-driven 4L60transmission dynamometer-cycling test used by General Motors. Initial friction tests often include a shift quality test run in a vehicle and vehicle shudder testing to test compatibility of the fluid with torque converter clutch operation.

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Automatic Transmission Fluids TABLE 2.4

2-13

SAE#2 Test Parameters for Various ATF Friction Durability Tests

Composition plate Grooving Test length (h) Test length (cycles) Total energy (J) Clutch release (kPa) Clutch pack clearance (mm) Motor speed (rpm) Clutch apply (kPa) Fluid temp (◦ C) Fluid volume (ml) Cycle rate (cycles/min) Plate configuration Energy per engagement (J/cm2 )

DEXRON®-IIIH flat plate

Ford MERCON® 15k flat plate

Ford MERCON® V 20k flat plate

SD1777 None 100 18,000 27,000 150 1.02 ± 0.13 3600 345 140 600 3 S-F-S-S-F-S 85.6

SD1777 Cross hatch 63 15,000 20,740 ≥75 0.70 ± 0.13 3600 275 ± 5 115 ± 3 305 ± 5 4 S-F-S-S-F-S 88

SD1777 Cross hatch 83 20,000 20,740 ≥75 0.70 ± 0.13 3600 275 ± 5 135 ± 2 305 ± 5 4 S-F-S-S-F-S 88

Antiwear is another critical aspect of ATF performance for which there are a multitude of tests. Ford’s MERCON® specification requires an ASTM D2882 vane pump test, while the MERCON® V specification requires the same vane pump tests and a FZG Gear Wear Test (ASTM D5182), a Four-Ball Wear Test (ASTM D-4172), a Falex EP test (ASTM D3233), and a Timken Wear Test (Modified ASTM D2782). All these additional tests have a requirement to be run at 150◦ C. General Motors in its recent DEXRON® specifications requires a special sprag (one-way) clutch wear test, and the modified ASTM D2882 vane pump test requiring 15 mg maximum weight loss. Antiwear is further assessed through the inspection of the dynamometer-cycling test, where it is examined for any unusual wear vs. reference on the gears, bearing, etc. Resistance of the ATF to oxidation is important for fill-for-life considerations. The GM DEXRON® specification employs a full-scale 4L60 transmission driven by an electric motor at a 163◦ C sump temperature and 450 h to assess oxidation resistance. Requirements include a change in total acid number (TAN) of less than 3.25 mg KOH/g fluid, a sludge amount less than or equal to a reference oil, and a change in infrared carbonyl absorbance of 0.45 max. Ford MERCON®V testing uses a special bench ABOT to assess oxidation for 300 h at 155◦ C with a change in TAN of 3.5 max, pentane insolubles of less than 0.35%, and other criteria being required. Corrosion ratings of copper and lead strips suspended in the ATF during the ABOT procedure are also requirements. The lead strip weight loss cannot exceed 3.0% and the visual rating of the copper strip at 50 h and at 300 h is 3b maximum. Copper corrosion protection is usually determined by conventional ASTM D-130 testing which is provided for in both MERCON® and DEXRON® specification requirements with a 1b maximum rating at 3 h and 150◦ C. Note that this is a significantly higher temperature than the 121◦ C temperature conventionally used for gear lubricants. The ASTM D665 Turbine Oil Rust test is employed in both MERCON® and DEXRON® specifications to measure the ATF’s ability to protect against rust. Additionally, the DEXRON® specifications require the ASTM D1748 Humidity Cabinet Rust Test. Corrosion is also tested for in other tests like the MERCON® ABOT where the catalytic copper strip can have only a 3b maximum rating. In DEXRON® oxidation testing in the 4L60 transmission parts are monitored for any corrosion that exceeds the reference. Good antifoaming performance is also a key requirement for ATFs. Tests for antifoam performance include the ASTM D892 three-sequence foam test used by most manufactures. Additionally, Ford adds a fourth high temperature sequence, the ASTM D6082 test. GM also has a special foam test and an aeration test to measure air-entrainment, and it also limits foam levels in its oxidation and cycling tests run in full-scale transmissions. Future trends in ATF testing will most probably concentrate on two areas of performance. First, continued improvement in friction stability and oxidation resistance will be important for fill-for-life

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Handbook of Lubrication and Tribology

performance. Stability related features such as viscosity at all temperatures are becoming more important to maintain new transmission performance for the life of the vehicle. The second driving force in testing is the introduction of new technologies. Dual clutch transmissions, for example, may require durability testing with new friction materials using longer high-energy engagements not typical of conventional automatic transmission shifting clutches. Push-belt CVTs like those developed by Van Doorne Transmissie and adapted by OEMs in Europe, Japan, and the United States require automatic transmission-type fluids with the capability of operating with conventional clutches and also requiring high metal–metal friction for low slippage and good torque transfer in the belt to variator interface. Conventional automatic transmissions and CVT transmissions may be fitted with start-up clutches replacing the torque converter for better fuel efficiency. This again will require specialized testing to model the stresses on the fluid in this sort of application.

References [1] Kemp, S.P. and Linden, J.L., “Physical and Chemical Properties of a Typical Automatic Transmission Fluid,” SAE Paper Number 902148, International Fuels and Lubricants Meeting and Exposition, Tulsa, Oklahoma, October 22–25, 1990. [2] Watts, R.F. and Szykowski, J.P., “Formulating Automatic Transmission Fluids with Improved Low Temperature Fluidity,” SAE Paper Number 902144, International Fuels and Lubricants Meeting and Exposition, Tulsa, Oklahoma, October 22–25, 1990. [3] Sprys, J.W., Vaught, D.R., and Stephens, E.L., “Shear Viscosities of Automatic Transmission Fluids,” SAE Paper Number 941885, Fuels and Lubricants Meeting and Exposition, Baltimore, Maryland, October 17–20, 1994. [4] Haviland, M.L., Anderson, R.L., Davison, E.D., Goodwin, M.C., and Osborne, R.E., “Dexron-II Automatic Transmission Fluid Performance,” SAE Paper Number 740053, Automotive Engineering Congress, Detroit, Michigan, February 25–March 1, 1974. [5] Ross, W.D. and Pearson, B.A., “ATF-TYPE F Keeps Pace with Fill-for-Life Requirements,” SAE Paper Number 680037, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [6] Kobe, R.A. and Wagner, J.C., “The Chrysler TorqueFlite and Automatic Transmission Fluid,” SAE Paper Number 680036, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [7] DaimlerChrysler Material Standard, MS-9602, Change E, 2002. [8] ZF Factory Fill ATF Standard ZFN 13014, 2000-12. [9] API 1509, “Engine Oil Licensing and Certification System,” 15th ed., April 2002, American Petroleum Institute, Washington, DC. [10] Yang, Y., Lam, R., and Fuji, T., “Prediction of Torque Response During the Engagement of Wet Friction Clutch,” SAE Paper Number 981097, International Congress and Exposition, Detroit, Michigan, February 23–26, 1998. [11] Ward, W., Snyder, J., Lann, P., and Derevjanik, T., “ATF Nylon Degradation,” SAE Paper Number 971625, May 1997, International Spring fuels and Lubricants Meeting, Dearborn, Michigan, May 5–8, 1997. [12] Coleman, L., “Development of Type F Automatic Transmission Fluids,” SAE Paper Number 680039, January 1968, Automotive Engineering Congress, Detroit, Michigan, January 8–12, 1968. [13] “Automatic Transmission Fluid,” Ford Motor Co., Manufacturing Standards. M2C33 E-F, March 1, 1967. [14] Ford Motor Company, “A Specification for MERCON® V,” Revised October 1, 1998. [15] Riga, A., Patterson, G., Pistillo, W., Scharf, C., and Ward, W., “Automatic Transmission Fluid Compatibility with Nylon Components by Thermomechanical Analysis and Thermogravimetry,” Thermochimica Acta, 226, 1993, 363–368. [16] Tipton, C. and Schiferl, E., “Fundamental Studies on ATF Friction I,” SAE Paper Number 971621, International Spring Fuels and Lubricants Meeting, Dearborn, Michigan, May 5–8, 1997.

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Automatic Transmission Fluids

2-15

[17] Johnson, M., Korcek, S., and Zinbo, M., “Inhibition of Oxidation by ZDTP and Ashless Antioxidants in the Presence of Hydroperoxides at 160◦ C — Part I,” SAE Paper Number 831684, Fuels and Lubricants Meeting, San Francisco, California, October 31–November 3, 1983. [18] Slough, C., Ohtani, H., Everson, M., and Melotik, D., “The Effect of Friction Modifiers on the LowSpeed Friction Characteristics of Automatic Transmission Fluids Observed with Scanning Force Microscopy,” SAE Paper Number 981099, February 1998. [19] Zhu, Y., Ohtani, H., Greenfield, M., Ruths, M., and Granick, S., “Modification of Boundary Lubrication by Oil-Soluble Friction Modifier Additives,” Tribology Letters, 15(2), 2003. [20] Ford Engineering Material Specification, “ESP-M2C138-CJ,” released September 1974. [21] Ford Engineering Material Specification, “ESP-M2C166-H,” released June 1981. [22] “A Specification for MERCON® Automatic Transmission Fluid Trademarked for Service in Vehicles Sold by The Ford Motor Company,” Revised April 1, 1992. [23] Florkowski, D. and King, T., Chrysler Corporation, Skroubul, A., Texaco Lubricants, and Sumiejski, J., Lubrizol Corporation, “Development and Introduction of Chrysler’s New Automatic Transmission Fluid,” SAE Paper Number 982674, International Fuels and Lubricants Meeting and Exposition, San Francisco, California, October 19–22, 1998.

© 2006 by Taylor & Francis Group, LLC

3 Rear Axle Lubrication 3.1 3.2 3.3

Rear Axle Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Viscosity Classifications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Gear Oil Classification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3-2 3-2 3-3

Service Designations in Current Use • Service Designations Not in Current Use

3.4 3.5

Performance Requirements for Gear Oils . . . . . . . . . . . . Gear Oil Composition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3-4 3-5

Base Oil • Viscosity Modifiers and Pour Point Depressants • Performance Package

3.6 3.7

Arup Gangopadhyay Ford Motor Company

Farrukh Qureshi The Lubrizol Corporation

Issues and Challenges for Rear Axle Fluids. . . . . . . . . . . Fuel-Efficient Gear Lubricants . . . . . . . . . . . . . . . . . . . . . . . .

3-8 3-8

Axle Efficiency Tests • Vehicle Tests • Spin Loss Tests • Effect of Gear Surface Finish on Efficiency • Effect of Cold Start on Axle Efficiency • Limited Slip Differentials

3.8 Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3-19 3-20

This chapter will briefly discuss axle hardware, viscosity classifications, fluid requirements, and lubricant composition. The primary focus will be on development of fuel-efficient lubricants while maintaining axle durability. An automotive engine develops maximum power at a relatively high speed. The torque of the engine is modified in various stages until it becomes a propulsive force at an appropriate speed at the interface of tires and the road. The rear axle is the final stage in the drive train of a vehicle, which is responsible for appropriately transforming the engine power into useful propulsive force. The power from the engine and transmission is transferred to the rear axle through the driveshaft. The driveshaft is connected to the pinion gear inside the axle housing, which is partly filled with lubricant. Rotation of the driveshaft turns the pinion gear, which in turn rotates a contacting ring gear, as shown in Figure 3.1. The ring gear is connected to two shafts on both sides which extend outside the axle housing and connect to the wheels. The pinion gear is supported on two roller bearings and the two shafts connecting to the wheels are supported on two axle roller bearings mounted on the housing. Therefore, energy losses in the rear axle are due to (1) losses in the four bearings, (2) shearing of rear axle lubricant in the axle housing, and (3) frictional loss in pinion and ring gear contacts. The amount of preload on the pinion bearing also contributes to frictional loss. An important component of the rear axle system is the gear oil which is required to play a critical role in the efficient and durable operation of the rear axle. Axle durability is related to the pinion bearings and also depends on the wear of ring and pinion gears. The bearing durability is in turn related to lubricant temperature. Excessive gear wear results in gear whine, which causes customer dissatisfaction.

3-1

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Handbook of Lubrication and Tribology

Ring gear

Pinion gear

FIGURE 3.1 A view of ring and pinion gears inside the axle housing.

3.1 Rear Axle Lubrication Usually spiral-bevel or hypoid gears are used in rear axles. Fluids used in rear axles are required to reduce wear, pitting, spalling, scoring, other types of gear tooth distress to increase the life and reduce the downtime of the equipment. Additional requirements may also include protection against oxidation, rust, copper corrosion, and foaming. Since vehicles have to operate in diverse climates, viscometrics at both high and low temperatures must also be tailored to provide adequate fluid film for protecting the surfaces. Most of these requirements, developed over the years, can be satisfied by passing industry specified standard tests. These tests have been shown to define lubricants with adequate properties. Additional requirements such as improving fuel economy require close cooperation between equipment manufacturers and lubricant formulators so that desired fluid properties are tailored to axle design. Several of the standard requirements and specifications are discussed in the following sections, followed by a detailed discussion on improving the efficiency and durability of rear axle fluids.

3.2 Viscosity Classifications Gear oil viscosity is the most important parameter that governs the fluid film thickness between operating surfaces and along with chemical additives, technology determines the degree of protection available for gears and bearings in the rear axle system. The viscosity of a fluid is its resistance to shear deformation and is usually measured by shearing the fluid under controlled temperature and shear rates. Another important parameter is viscosity dependence on temperature, as represented by the viscosity index of fluid. A detailed description of standard test methods will not be included in this work, but references will be provided as appropriate. Fluid viscosity also depends on shear rate and ambient pressure. Automotive gear lubricant viscosities are defined in SAE J306. The SAE gear oil viscosity classifications are shown in Table 3.1. SAE J306 has been updated in October 2005 and new viscosity grades have been added. These designations are used to specify the viscosity requirements for manual transmission lubricants. Multigrade gear oils that can maintain their film-forming characteristics over a large temperature range are being increasingly used. In order to reduce the temperature dependence of viscosity these multigrade gear oils contain significant amount of polymers. However, polymers are also susceptible to shear degradation under service conditions. If these polymers shear in service, it will lead to a drop in viscosity, resulting in reduced film thickness, and may eventually lead to equipment failure. It is required that the lubricant remain in its viscosity grade

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Rear Axle Lubrication

3-3

TABLE 3.1 Automotive Gear Lubricant Viscosity Classification SAE viscosity grade 70W 75W 80W 85W 80 85 90 110 140 190 250

Max temperature for viscosity of 150,000 cP (◦ C)a,b

Kinematic viscosity at 100◦ C, cStc Minimumd

Kinematic viscosity at 100◦ C, cStc Maximum

−55e −40 −26 −12 — — — — — — —

4.1 4.1 7.0 11.0 7.0 11.0 13.5 18.5 24.0 32.5 41.0

— — — — 90, 1.2 ; otherwise it denotes 1 . Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.

Another detail to be noted is that although C is the machined clearance in all three lobes; when assembled, this dimension does not physically appear in the bearing. The largest concentric clearance, which occurs at the junction of the three lobes, is less than the machined clearance C, and will here be denoted by CM . It is given by C¯ M = (CM /C ) = (1 − m /2)

(7.30)

whereas C¯ m , as before, is given by (1 − m ). Figure 7.25 shows the locus of shaft center in a 3-lobe bearing of 100◦ arc extent. Since, as seen here, the shaft never rises above the horizontal centerline, the minimum film thickness will always be located in the bottom pad, and is given by hmin (1 − 1 )

(7.31)

Table 7.9 and Figures 7.25 and Figure 7.26 give the performance characteristics of the 3-lobe bearing for three (L /D) ratios and two ellipticity ratios. Together with the 3-groove circular bearing given in Table 7.7, which represents the case m = 0, a three point variation in m is provided from which performance for intermediate values of m can be obtained by cross plotting [25].

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Handbook of Lubrication and Tribology 2.4 2.0

1 2 3 4

3 1

1.6

4

L/D = 1/4 L/D = 1/2 L/D = 1 L/D = 1 + 1/2

m=0

2

1.2 0.8 2.6 3 2.2

Power loss factor H

1.8

1

4

m = 0.25 2

1.4 1.0 2.0

1

1.8

2

1.6

m = 0.5

3 1.4

4

1.2

2.2 1 2.1

4

m = 0.75 2

2.0 3 0.01

FIGURE 7.23

0.1 1.0 (mN/P) (R/C)3

10.0

Power loss in elliptical bearings.

It was said previously that the 3-lobe bearing is usually chosen for its superior stability characteristics and following is a brief illustration of its features vis-a-vis the elliptical bearing. Figure 7.27 shows the stable and unstable regimes plotted against Sommerfeld number for the two bearing types. Each of the constant lines gives the locus of the operation of a bearing with a fixed geometry as its speed is varied. The bearing parameter, η, is given by η=

µLD 2π W

 2   R W 0.5 C CM

which is independent of rotor speed and describes a certain bearing geometry. As speed is increased, the bearing will, via S, proceed along a line of constant η and eventually enter the unstable region. The

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Principles of Gas Turbine Bearing Lubrication and Design

7-35

f1 3

O1

3

CM

2 m

e1

O1 O

d

f2

O

m

R+C

O2 R

2

O3

Cm

C

e O9

e2

O2

m

f

O3 e3

1

f3

1 e1,2,3 =(e1,2,3/C) e =(e/C) m =(d/C)

FIGURE 7.24 Geometry and nomeclature of 3-lobe bearings.

W

v 0

90°

80°

0.2

70°

0.4

f 1/3 60°

0.6

2/3 50°

0.8

40°

m=0 30°

1.0

10°

Clearance boundary

20°

FIGURE 7.25 Locus of shaft center for 3-lobe bearings. (Taken from Pinkus, O., Trans. ASME, 78, 1956, 965–973. With permission.)

parameter η is most sensitive to bearing diameter and lobe clearance and less so to the length, L, and viscosity, µ. A move to a higher value of η, that is, to a more stable region is accomplished by either increasing µ, increasing L, increasing D, or decreasing C. Because the ordinate parameter ω(MC/W )0.5 also depends upon the clearance, changes in C will follow the slightly inclined dashed lines — toward the left if C is increased, toward the right if C is reduced. At light loads and small clearances, the 3-lobe

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Handbook of Lubrication and Tribology

Table 7.9

Performance of Three-Lobe Bearings. m = 1/2

m = 2/3

L /D

m

φ

1

S

Q

φ

1

S

1

0.2 0.4 0.6 0.8 1.0

42 53 55 50 30

0.58 0.63 0.71 0.815 0.965

0.45 0.18 0.10 0.048 0.0063

0.125 0.185 0.20 0.235 0.28

50 50 50 45 40

0.71 0.75 0.81 0.85 0.945

0.21 0.12 0.071 0.039 0.0095

0.18 0.185 0.20 0.21 0.23

1/2

0.2 0.4 0.6 0.8 1.0

45 55 55 50 30

0.57 0.63 0.71 0.815 0.905

0.84 0.40 0.20 0.084 0.011

0.265 0.31 0.335 0.425 0.50

50 52 50 45 40

0.71 0.76 0.81 0.86 0.945

0.43 0.21 0.11 0.054 0.0125

0.355 0.35 0.38 0.41 0.465

1

0.2 0.4 0.6 0.8 1.0

45 45 45 40 30

0.575 0.65 0.745 0.845 0.965

2.5 1.0 0.41 0.13 0.021

0.37 0.51 0.54 0.61 0.75

62 55 58 52 44

0.70 0.75 0.80 0.86 0.945

1.16 0.59 0.33 0.12 0.025

0.53 0.53 0.53 0.574 0.58

Q

Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.

b = 100°

bG = 20°

– Power loss, H

3

2

m = 2/3

m = 1/2 m=0 1 0

0.2

0.4

0.6

0.8

1.0

em

FIGURE 7.26

Power loss in three-lobe bearings.

bearing is better than the elliptical bearing. Under heavier loads, the elliptical design is the more stable bearing. It should, however, be kept in mind that it is precisely the low load range that is the troublesome region of bearing stability. When a bearing is absolutely stable, the whirl ratio approaches zero and Figure 7.27(b) shows the variation of the whirl ratio at the threshold of instability. The whirl ratios with the 3-lobe and the elliptical

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design (b) 01

(a) 103 0. 00

h ig ed e sp

0.7 Elliptical

0.6

1=

0. 00

1

Unstable

3-lobe

Half-frequency whirl

0.5 0.4

3-lobe

0.3 0.2

1=

0. 01

10 L arge clear ance

Whirl ratio, (vj/v)

H

1=

v(MRC/W)1/2

102

7-37

Stable Low speed 1 0.01 0.1

Elliptical

(mNLD/W)

1.0

Small clear ance 10.0

(R/C)2

3-lobe b = 100° m = 0.5 Elliptical b = 160°h = (mLD/2pW)(R/C)2(W/CM3)1/2

FIGURE 7.27 frequency.

0.1 0 0.01

0.1

1.0

10.0

(mNLD/W) (R/C)2 3-lobe b = 100° m = 0.5 Elliptical b = 160°

Stability characteristics of elliptical and 3-lobe bearings (a) Stability regimes. (b) Instability threshold

bearings share similar characteristics. They both rise sharply at S = 0.1 to 0.2 and remain fairly constant thereafter. The elliptical bearing which is the less desirable for lightly loaded applications is seen to have a whirl ratio in excess of 0.5, thus a worse ratio than the 3-lobe bearing. The whirl ratio is an important parameter in determining the stability threshold for a flexible rotor which is always lower than that for a rigid rotor. Thus, Figure 7.27(b) can be viewed as the highest possible stability that can be achieved with these bearings [23]. 7.4.2.4 Tilting Pad Bearings Unlike the previously considered designs, the tilting-pad bearing is a generic name that covers many permutations. Its primary characteristic is that the individual pads are not fixed in position, but are pivotsupported so that during operation not only does the shaft move in response to operational conditions, but so do the pads, and each pad in a different fashion. A general picture of a pivoted shoe bearing is shown in Figure 7.28. The complexity of the design is partly evidenced in the configuration of a single pad given in Figure 7.28(b). Several things ought to be noticed. In the first place, the criterion of hmin as a measure of load capacity loses its meaning somewhat here since this hmin is not a fixed distance; however, the film thickness over the pivot, hp , is a geometrically fixed point. Under excessive load, given that the pad at hmin can yield whereas at hp it cannot, failure is more likely to occur at hp . Thus, the critical quantity here is perhaps hp rather than hmin . The next thing to realize is that the center of curvature of the pad is not fixed in space; when the pad rocks above the pivot, its center of curvature moves either in a positive or negative angular direction, shown in Figure 7.28(b) by ±γ . Next, should the preload be too low, some of the pads on the top of the bearing may become unloaded, in which case, as shown in Figure 7.29, the fluid film frictional moment about the pivot will make the leading edge of the pad scrape against the journal and cause“flutter,” obviously an undesirable contingency. The condition that the top pads not be unloaded is dictated by the amount of preload and shaft position; the lower the preload and the higher the , it is more likely that some of the pads will become unloaded. Figure 7.30 gives a sample graph, in terms of m and m, when a pad is likely to be unloaded. In this respect, loading between the pads, when the journal may reach m > 1, is a more undesirable mode of operation [30].

© 2006 by Taylor & Francis Group, LLC

7-38

Handbook of Lubrication and Tribology (a)

Pivot Pivot clearance circle

Pad R-Cm e w

R

Shaft

b1

(b)

W

Cm

Concentric journal

Pad 1

us1

e1 hp1

bs1

b1

b

N

E O1 Oe

Pad 2

1

O

R

up1 hmin 1 Concentric pad uE1

FIGURE 7.28

Tilting-pad journal bearing. Three-pad tilting-pad bearing. Geometry of tilting pad.

The number of possible design parameters and operating modes in a tilting-pad bearing is large. Some of these design options are: • • • • • • •

Number of pads, 3 > n > 8 Angular extent of pads, including the option of a variation of β for various pads Load vector passing between pads or over a pad Central or eccentric pivot location, that is, a choice of value for (βp /β) (L /D ) ratio Preload, m = 1 − (Cm /C ) Pad inertia, which often determines the ability of the pad to follow or track journal motion.

7.4.3 Thrust Bearings Much of what has been said previously about journal bearings applies also to the behavior of thrust bearings. It thus remains only to point out the differences that arise due to the different geometry of a thrust bearing shown in its generic elements in Figure 7.18[a].

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

7-39

Frictional forces

Frictional moment

N O + Shaft center

FIGURE 7.29 Unloaded tilting pad.

0.6

0 0.3.4

2

4

0.

1

0.

0

0.

0.6

em

1.0 0.8

em All pads always loaded

1.2 1.0 0.8

For the given value of m maximum em at which all pads are loaded.

2

Load

0

0.4 0.3 0. 2 0. 1

0.6 0.5

0.7

m

0.8

0

0.

Load

Some pads always unloaded

FIGURE 7.30 Regime of unloaded pads in a 5-pad tilting pad bearing.

© 2006 by Taylor & Francis Group, LLC

7-40

Handbook of Lubrication and Tribology

Thrust bearings are simpler to handle in that no cavitation occurs and in that it is sufficient to solve only for one pad (for a parallel runner all pads are identical). The geometry of the film, on the other hand, is more varied. The film shape in journal bearings is more or less universal, namely that prevailing between two eccentric circular cylinders. In thrust bearings it can be anything — shapes with one or two directional tapers, with or without flats, crowned profiles, pocket bearings, and finally tilting-pad designs. Another simplification with thrust bearings is that no instability problems, such as those that occur with journal bearings, arise in their operation. There is, thus, no need to evaluate stiffness and damping. The Reynolds equation for thrust bearings has to be written in polar instead of rectangular coordinates. In parallel to journal bearings, turbulence is accounted for on a point-by-point basis, here a function of r as well as θ ∂ ∂r



3

rh Gz µ




∂p ∂r



 3


1 ∂ h Gx ∂p r ∂h + · =6 r ∂θ µ ∂θ (L/R2 )2 ∂θ

(7.32)

where Gx , Gz are the turbulence coefficients in the θ and r directions respectively, both functions of the Reynolds number given by Re = ρrωh/µ = f (r, θ ) The expressions above differ from those for the journal bearing in that they have a dependence on r. This is due to the variation of the Couette flow with r and to the film thickness being a function of both coordinates, r and θ. 7.4.3.1 Tapered Land Bearings The simplest tapered land bearing is one which has a constant angular taper, or h(θ) − h2 + δθ (1 − θ/β)δθ = (h2 − h1 )

(7.33)

with its geometry as shown in Figure 7.31(a). This equation, independent of r, is valid for a bearing surface with a circumferential taper alone. As will be shown later, the exact shape of the fluid film between fixed

B A

A L

Load U

Runner

h1

du B

Cut A–A

FIGURE 7.31

Tapered land bearing.

© 2006 by Taylor & Francis Group, LLC

h2

Principles of Gas Turbine Bearing Lubrication and Design

7-41

values of h1 and h2 does not affect the results appreciably. Thus, by their simplicity, the on-dimensional taper solutions provide a useful key for evaluating the bearings in general. The several crucial parameters in journal bearings are β. (L /D ), and (e /C ). Parallel quantities appear in thrust bearings, namely, β, the angular extent of the pad; (L /R2 ); and (h2 /δθ ) with δθ (like C) being a geometric quantity and h2 being the trailing film thickness at which the bearing is run. It should be also noted here that hmin = h2

(7.34)

Solutions for the tapered land bearing are given in Table 7.10, where Q r = (Qr /π R2 NL δθ )

(7.35)

is the side leakage with the index R1 indicating the leakage along the inner radius and R2 indicating the leakage along the outer radius. The total side leakage is then Q r = [Qr |R1 + Q r |R2 ]π R2 NL δθ The leakage out the end of the pad, Q2 , is given by: Q2 = 0.5π NLh1 (R1 + R2 ) + Q2 p π R2 NL δθ

(7.36)

where the first right-hand term is the shear flow and does not involve any computer obtained coefficients. The value of Q 2p can be obtained from Table 7.10 by subtracting Q r from (Q r + Q 2p ). Table 7.11 shows the relative load capacities and friction of three different thrust bearing configurations. One is a plane slider, that is, an inclined rectangular block; the second, a slider with an exponential film profile; and the third is the tapered land geometry of Equation (7.33). As seen, the results for a given value of (h1 /h2 ) are nearly identical, confirming our assertion that once h1 and h2 are fixed, the exact variation in h between these values is not of great importance. In all of the above results, it should be noted that P is the unit pressure given by: P = W /Area = 360 WT /[n βπ L (R2 + R1 )] where β is in degrees. WT is the total load on the thrust bearing and n the number of pads. Also it should be noted that the data for flow and power loss in Table 7.10 are for a single pad so that the total flow and losses are QT = n Qpad

HT = n Hpad

7.4.3.2 Composite Tapered Land Bearings A more practical and preferred thrust bearing geometry is a tapered land bearing having tapers in both the circumferential and radial directions with a flat portion at the end of the film. Its advantages are (1) it has higher load capacity, (2) has lower side leakage and, (3) at low speed and during starts and stops it provides a flat surface for supporting the load, thus minimizing wear. The geometry of such a bearing is shown in Figure 7.32. Its film thickness is given by  h = h11 −

   h11 − h12 h11 − [h11 − h12 /L](R − r1 ) − h2 (r − R1 ) − θ L bβ for 0 < θ < bβ h = h2 , constant for bβ < θ , β

© 2006 by Taylor & Francis Group, LLC

(7.37)

7-42

Handbook of Lubrication and Tribology

Table 7.10

L/R2 1/3

Solutions for Tapered Land Thrust Bearings

b1 /Sθ 1

1/2

1/4

1/8

1/2

1

1/2

1/4

3/8

2/3

1

1/2

1/4

1/8

β (deg)

µN p




L δθ

2

r = (r − R1 )/Lθ = θ/β Center of pressure

Qr

H δθ

at R1

at R2

Q r + Q 2p

θ

r

π µN 2 R24

80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30

1.423 1.180 0.947 0.870 0.321 0.257 0.225 0.211 0.0855 0.714 0.0652 0.0635 0.0278 0.0247 0.0238 0.0242

0.34 0.32 0.28 0.235 0.35 0.32 0.29 0.245 0.35 0.32 0.29 0.235 0.36 0.33 0.29 0.25

0.40 0.44 0.81 0.75 0.47 0.44 0.40 0.36 0.47 0.44 0.41 0.36 0.48 0.45 0.41 0.37

0.87 0.84 0.81 0.75 0.87 0.84 0.79 0.74 0.87 0.83 0.78 0.70 0.85 0.81 0.75 0.67

0.64 0.025 0.61 0.605 0.71 0.69 0.67 0.66 0.78 0.76 0.74 0.73 0.83 0.815 0.795 0.78

0.37 0.45 0.49 0.51 0.37 0.47 0.50 0.51 0.41 0.45 0.505 0.52 0.465 0.50 0.51 0.565

2.44 1.685 1.26 0.95 3.94 2.70 2.00 1.57 5.96 4.25 3.23 2.54 8.51 6.23 4.88 3.91

80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30

1.72 1.494 1.435 1.489 0.402 0.3585 0.352 0.370 0.1138 0.1062 0.1080 0.1103 0.0402 0.0399 0.0423 0.0470

0.23 0.19 0.145 0.11 0.23 0.19 0.15 0.11 0.24 0.20 0.15 0.11 0.25 0.20 0.16 0.11

0.405 0.36 0.31 0.20 0.41 0.33 0.31 0.26 0.42 0.27 0.32 0.27 0.42 0.28 0.32 0.27

0.75 0.69 0.61 0.57 0.74 0.61 0.60 0.53 0.72 0.65 0.56 0.49 0.70 0.62 0.53 0.44

0.62 0.61 0.60 0.59 0.685 0.67 0.655 0.65 0.755 0.735 0.72 0.71 0.81 0.78 0.77 0.765

0.48 0.51 0.53 0.55 0.46 0.52 0.53 0.55 0.48 0.52 0.54 0.56 0.50 0.53 0.55 0.57

2.90 1.96 1.47 1.13 4.72 3.33 2.49 1.92 7.32 5.29 4.065 3.18 10.81 8.06 6.30 5.01

80 55 40 30 80 55 40 30 80 55 40 30 80 55 40 30

2.240 2.185 2.320 2.590 0.538 0.537 0.578 0.653 0.1598 0.1655 0.1820 0.2035 0.0599 0.0649 0.0737 0.0861

0.12 0.082 0.052 0.033 0.13 0.084 0.0535 0.034 0.13 0.087 0.055 0.035 0.14 0.09 0.056 0.036

0.35 0.295 0.245 0.200 0.35 0.30 0.25 0.20 0.36 0.30 0.25 0.21 0.365 0.31 0.25 0.21

0.60 0.53 0.48 0.44 0.58 0.51 0.45 0.40 0.56 0.46 0.40 0.38 0.53 0.44 0.35 0.29

0.61 0.60 0.59 0.59 0.67 0.66 0.65 0.645 0.735 0.72 0.71 0.705 0.79 0.78 0.765 0.75

0.50 0.55 0.58 0.61 0.51 0.56 0.59 0.61 0.53 0.57 0.60 0.62 0.55 0.58 0.61 0.63

3.06 2.12 1.57 1.20 5.07 3.59 2.70 2.07 8.00 5.70 4.43 3.46 12.07 8.98 6.94 5.47

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Principles of Gas Turbine Bearing Lubrication and Design

7-43

Table 7.11 Performance of Thrust Bearings With Various Film Configurations Plane slidera

α

Exponential sliderb PL 2 h22

P= 2.00 2.50 2.85

µωR24

0.0810 0.113 0.135

0.0819 0.1137 0.435 F=

2.00 2.50 3.04

Sector padc

0.0826 0.106 0.125

Fh2 µω24 R24

0.66 0.74 0.84

0.81 0.875 0.95

0.78 0.825 0.88

Note: a h = αx b h = k1 e k2 x c h = h2 + δ(1 − θ/β) Source: Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.

h12 u

(r,u)

h2

v

Ro

Fla

t

h11

L

bb

Ri

b v Flat at R1

2h2∗ h12

h11

h2

d21

Ro

Ro Ri

Ri

FIGURE 7.32 Composite tapered land bearing.

Normalizing all h’s by h2 and all radii by R2 , we have 
  θ θ r = (L/R2 ) − 1 h − h 11 − (h − 1) − δr 1− bβ (L/R2 ) bβ for 0 < θ < bβ h = 1, for bβ < θ < β

© 2006 by Taylor & Francis Group, LLC

(7.38)

7-44

Handbook of Lubrication and Tribology Table 7.12 Rei 500

1500

3500

Composite Tapered Land Thrust Bearings Q/R12 h2 ω

Wh22

Hca h2

µR14 ω

µR14 ω2

at θ = 0

at R1

at R2

H W ωh2

0.192 0.182 0.175 0.168 0.151 0.337 0.307 0.289 0.275 0.240 0.567 0.499 0.463 0.435 0.369

3.92 3.88 3.78 3.68 3.42 8.45 7.93 7.59 7.31 6.63 15.6 14.0 13.2 1.25 11.2

1.58 1.58 1.59 1.59 1.60 1.61 1.62 1.62 1.63 1.64 1.63 1.64 1.64 1.65 1.66

0.296 0.294 0.293 0.292 0.290 0.324 0.321 0.319 0.318 0.314 0.339 0.334 0.331 0.329 0.325

0.437 0.445 0.451 0.456 0.469 0.455 0.465 0.471 0.476 0.490 0.462 0.475 0.482 0.488 0.504

17.6 18.3 18.5 18.7 19.1 21.5 21.6 21.8 21.9 22.3 23.2 23.2 23.3 23.4 23.7

Note: (R2 /R1 ) = 2; β = 40◦ , h11 = 3, δr = 0.5; b = 0.8 a H includes losses over a 10◦ oil groove. All results are per individual pad. c

The expression for h has, as seen, three arbitrary parameters: • h 11 = (h11 /h2 ) — the dimensionless maximum film thickness at the lower left corner. • δ r = (h11 − h12 )/h2 — the radial taper along the leading edge θ = 0. • b = the friction of β tapered. In an optimization study in which both load capacity and lower power losses were considered, the following desirable proportions for the above three parameters were arrived at: h = 3.0 δ r = 0.5

b = 0.8

Physically, the above numbers imply a maximum film thickness at (R1 , 0) of three times the one over the flat; an outward decrease in film thickness at the leading edge half that of h2 ; and a flat portion equal to 20% of the pad’s angular extent. The performance of such a bearing for the case of a 40◦ bearing pad an (OD/ID) ratio of 2 is given in Table 7.12. The table provides data for both turbulent and laminar operation. The following comments will, perhaps, be useful: • The values of the Reynolds number Re1 = ρR1 ωh2 /µ1 • The losses as represented by Hc in column four include the losses over a 10◦ oil groove; the losses H over the pad above can be obtained from the last column in Table 7.12. • The flow QIN represents the inflow at θ = 0. The outflow will, of course, be given by Q2 = QIN − (QR1 + QR2 ) • The lowest value of Re1 given is 500. This value is close to laminar operation. • For the total bearing, the values of W , Hc , Q, and H should all be multiplied by the number of pads.

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Principles of Gas Turbine Bearing Lubrication and Design

7-45

v

(a)

Pressure profile U

(b) Resultant v Original shape Thermally and elastically distorted pad

FIGURE 7.33 The hydrodynamics of a tilting pad thrust bearing. (a) Zero load capacity of centrally pivoted slider. (b) Generation of a hydrodynamic film due to thermal and elastic distortions.

7.4.3.3 Tilting Pad Bearings The comments made about the tilting pad journal bearing regarding its complexity and large number of parameters apply equally well to the thrust bearing. However, in the case of a pivoted thrust pad such as the one shown in Figure 7.33, an additional complication overshadows the other difficulties; that there is theoretically no solution to a planar centrally pivoted sector. This can be deduced from the pressure profile sketched in Figure 7.33(a). Such a profile must always be asymmetrical with respect to the center of the pad; an asymmetrical pressure profile would impose a moment about the pivot tending to align the pad parallel to the runner. However, a parallel pad produces no hydrodynamic pressures, thus making the working of such an arrangement impossible. Yet such centrally pivoted, plan surface thrust bearings are widely used and they perform exceedingly well. Various theories have been advanced and stratagems employed to explain the workings of these bearings and to obtain a solution to the problems. Among these are: • Thermal or density wedge — The variation in viscosity or density of the oil is often credited with generating hydrodynamic forces in the parallel film. At best, such effects produce forces which come nowhere near the heavy loading supported by such bearings. • Thermal and elastic distortion of the pad — As shown in Figure 7.33(b), thermal and elastic stresses may crown a pad, so that in essence it produces a convergent–divergent film. In that case, it is possible for the resultant load to pass through the pivot and the pad can support a load. However, such bending can occur either with very thin pads or extremely high-temperature gradients. Yet such bearings performance satisfactorily even with very thick pads and under conditions of minimal heat generation. • Incidental effects — There are a number of incidental features which may play a more important role than the above theoretical explanations. Among these are: (a) Machining inaccuracies on the faces of both runner and bearing and rounded off edge at entrance to the pad, which in effect constitute a built-in taper.

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7-46

Handbook of Lubrication and Tribology

(b) Misalignment between runner and pads during assembly or during operation. (c) Pivot location not exactly at 50% of pad angular extent. These factors would combine to generate hydrodynamic forces and they are perhaps the most likely explanation for the satisfactory working of tilting pad thrust bearing.

7.5 Low-Speed Bearings One of the requirements in the bearing described in the previous section is a proper lubrication system. This includes a pump delivering oil at 10 to 50 psi supply pressure with all the accompanying equipment such as oil tank, filters, piping, sump, and cooling arrangements. When the bearings run at relatively low speed involving low power dissipation and therefore low bearing temperatures — as in fans, blowers, and some compressors — one can simplify the system by employing oil-ring lubrication. This consists of a self-contained oil delivery package placed adjacent to the bearing which dispenses with all the auxiliary equipment required for a more demanding operation. Figure 7.34 shows the components of an oil-ring lubrication setup. The ring, riding on the top of the exposed shaft, is a sort of viscous drag device that lifts oil from the sump and deposits it on the shaft. It is clear that in comparison to a pressurized supply system where the oil is distributed along an axial groove, the amount of oil lifted is not sufficient to provide the bearing with a complete oil film, and therefore an important parameter in oil ring operation is the amount of oil the bearing receives relative to what it needs for a full film. This is called the starvation ratio and is given by ˆ z = Qz /QzF Q where Qz is the side leakage under starved conditions and QzF is the side leakage for a full film. Thrust bearing

Shaft

vring v

Journal bearing Oil flow

Sump oil

FIGURE 7.34

Oil-ring lubrication system.

© 2006 by Taylor & Francis Group, LLC

Oil ring

Cooling water

Sump

Principles of Gas Turbine Bearing Lubrication and Design

7-47

7.5.1 Regimes of Operation The value of starvation ratio depends on shaft and ring speeds. Due to the centrifugal effects of the rotating ring and attached oil one can distinguish four regimes of ring behavior, as shown generically in Figure 7.35. The characteristics of these regimes are as follows [15]: Regime I : At the low end of journal rotation there is contact between ring inner surface and the journal, and the linear speeds of the two mating surfaces are about the same. There is thus a linear rise in ring rpm with the rpm of the journal. The oil delivered by the ring rises throughout this regime. At the upper end of Regime l, ring speed and oil delivery reach a local maximum. Regime II: At the beginning of this regime, direct frictional drag yields to a state of boundary lubrication between ring and journal. Due to this, slippage occurs and ring speed drops. Since the speed has decreased, so too does the amount of oil delivered by the ring. However, with further rise in journal speed, a full hydrodynamic film is established between journal and ring. The reduced viscous friction (the friction coefficient may drop from 0.1 to 0.01) and the larger film between the mating surfaces bring about a rapid increase in both ring speed and oil delivery. Once again, at the upper end, a local maximum in ring speed is achieved. Oil flow, however, at the end of this regime is an absolute maximum and represents the highest possible oil delivery by the ring. Regime III: The drop in ring speed and oil delivery following Regime II is associated with the onset of ring oscillations in the plane of rotation. While small oscillations already appear during the trailing portion of Regime II, the values of σθ become, within a short span, very large and bring about a drastic reduction in oil delivery. While, due to these planar oscillations, the ring speed drops only slightly, the oil delivery is affected to such a point that at the end of this regime it approaches asymptotically zero.

QZF

NR

Regime I

Regime III

Regime IV

QR Regime II

sZ su

0 Journal speed

FIGURE 7.35 Oil ring behavior as function of journal speed.

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7-48

Handbook of Lubrication and Tribology u=0 u

W

N 0 eF 0°F e

u1F

u2F

0° f

fF u0F

u1 u2 u0 Extent of full film Extent of starved film

FIGURE 7.36

Full and starved fluid films.

The angular swing of the ring at its maximum values can be of the order of ±5 to 10◦ with an oscillatory frequency equal to that of ring rotational frequency. Regime IV : This regime, essentially beyond our interest, is characterized by conical and translatory vibrations of the ring. While the oscillatory motion abates and the ring speed once again tends to increase with journal speed, oil delivery is essentially zero. The frequency of both the conical and translatory (or axial) vibrations is that of ring rotational frequency. Starting with the oscillatory vibrations and proceeding through the two other modes of instability, the violent motion of the ring causes splash and a throw-off of oil from the surface of the ring, and partly also from the journal, so that little oil reaches the bearing. Figure 7.36 presents the hydrodynamics of a starved journal bearing vis-a-vis a full film or flooded condition. For a fixed load, speed, and supply oil temperature, the deviations of a starved bearing from one operating with a full fluid film are as follows: • The film starts later and terminates earlier, that is θ1 > θIF and θ2 < θLf , producing in essence something similar to a partial bearing, though the upstream, and sometimes also the downstream, boundary conditions are different. • The eccentricity increases, producing smaller values of hmin . • The attitude angle decreases, yielding a more vertical locus of shaft center. • Since oil supply is equivalent to side leakage, there is a reduction of bearing side leakage and consequently an increase in fluid film temperatures. This is somewhat mitigated by a reduction in power loss due to a shorter extent of the fluid film. Table 7.13 gives a set of solutions for a wide range of loads and levels of starvation [15]. It is seen that the effects of starvation are much more pronounced at light and moderate loads, which are due to the fact that, at heavy loads, the full film extends over a narrower arc and the pressure gradients at θ1 being higher, a lower amount of Q1 is required to form a film. The loci of shaft center, both for constant oil supply values Q1 and for constant loads, W , are given in Figure 7.37. It is seen that starvation displaces

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design Table 7.13

7-49

Theoretical Performance of Starved Journal Bearings

W

Qz %

C

φ, degs

Q1

Qz

θ1 , deg

βs , degs

0.491

0 0.9 3.1 12.0 29.0 53.0 100a

1.00 0.903 0.808 0.625 0.446 0.272 0.081

0 3 6 11 18 28 78

0. 0.100 0.200 0.400 0.600 0.800 1.03

0. 0.899 × 10−3 0.309 × 10−2 0.0122 0.0289 0.0534 0.100

180 175 171 163 155 144 105

0 13 22 40 60 86 150

1.965

0. 0.7 2.8 12.4 30.3 50.2 100a

1.00 0.906 0.818 0.657 0.516 0.395 0.285

0 5 9 17 26 37 57

0 0.100 0.200 0.40 0.600 0.800 1.03

0 0.208 × 10−2 0.0817 0.0357 0.0875 0.162 0.289

180 172 167 155 144 129 105

0 16 29 57 82 101 150

9.825

0. 1.5 6.4 15.6 28.0 60.0 100a

1.00 0.914 0.846 0.793 0.753 0.702 0.672

0 8 14 19 24 31 37

0. 0.100 0.200 0.300 0.400 0.600 0.809

0. 0.659 × 10−2 0.0291 0.0697 0.125 0.267 0.445

100 168 158 148 139 122 105

0 25 46 69 82 108 132

34.39

0. 4.8 19.4 38.7 61.3 100a

1.00 0.929 0.896 0.881 0.073 0.866

0 11 17 21 23 26

0 0.100 0.200 0.300 0.400 0.560

0 0.0192 0.0802 0.162 0.253 0413

180 160 144 131 120 105

0 38 73 79 86 110

98.25

0 12.9 38.7 66.1 93.8 100a

1.0 0.954 0.947 0.946 0.945 0.944

0 12 16 17 18 18

0. 0.100 0.200 0.300 0.400 0.423

0 0.0463 0.138 0.235 0.335 0.357

180 149 131 118 107 105

0 50 70 87 100 101

Note: µ = const; β = 150◦ ; (L/D) = 0.93 a Full fluid film. Source: Heshmat, H. and Pinkus, O. “Performance of Starved Journal Bearings with Oil Ring Lubrication,” J. Trib. Trans. ASME, 107, 1985, 23–32.

the locus of shaft center inward of the full film line, that is, toward higher eccentricities and lower values of attitude angle φ. From the above and the parametric studies described in Section 7.8 the following conclusions can be drawn: • Except at very low speeds, most oil ring bearings operate under starved conditions. • The load capacity of oil ring bearings, first increases then decreases with rising shaft speed. • The locus of shaft center of oil ring bearings is much closer to the vertical axis than in full film bearings. • An optimum in the (L/D) ratio exists in oil ring bearings which ranges from 0.6 to 0.8. • The effects of starvation are much more pronounced at low and intermediate loads than at high loadings.

© 2006 by Taylor & Francis Group, LLC

7-50

Handbook of Lubrication and Tribology v

0

90°

0.1 80°

0.2 W = 0.49

0.3

1.96

0.4

70°

Full film

Q1 = 0.6

0.5

60°

0.6

0.4 50°

0.7

9.82

0.8 0.9 1.0

f

0.2

40°

34.4

0.1

30° 20°

0°

Q1 = Const W = Const

10°

FIGURE 7.37 Locus of shaft center at different levels of starvation. (Taken from Heshmat, H., and Pinkus, O. J. Trib. Trans. ASME 107, 1985, 23–32. With permission.)

7.6 High-Speed and High-Temperature Oil-Free Bearings This section will consider bearings suitable for operation at extreme speed and temperature ranges. These two parameters may occur either together or independently, that is, although usually high speed implies high temperatures, the reverse is not always so. One can have lower or moderate velocities but the environment may be such that the bearing will be exposed to high temperatures and the fluid film, the lubricant, and bearing materials must be able to cope with it. The other important consideration in high-speed bearings is that of stability. As will be seen below, the likelihood of baring instability, known as half-frequency whirl, rises with rotational speed. This becomes particularly intense when speeds twice the natural frequency of the system are reached. The kinds of bearings that are possible candidates for such applications are gas bearings, either hydrodynamic or hydrostatic, compliant surface geometries, and magnetic bearings.

7.6.1 Gas Bearings The differential equation governing the behavior of gas bearings is that given by Equation (7.5). For liquids ρ is a constant and the density terms fall out of the equation. In gas bearings it varies with both pressure and temperature. In most cases the perfect gas equation is applicable, or P = RT ρ

© 2006 by Taylor & Francis Group, LLC

(7.39)

Principles of Gas Turbine Bearing Lubrication and Design

7-51

Two factors make the gas film in bearings isothermal; one is the low heat generation; the other is the high thermal capacity of the bearing shell as compared to the tiny volume of gas in the film. We therefore have ρ = RT p = const · p

(7.40)

and Equation (7.39) becomes ∂ ∂x



ph 3 µ




∂P ∂x

 +

∂ ∂Z



ph 3 µ




∂p ∂z

 = 6U

∂(ph ) ∂x

(7.41)

All the solutions given subsequently are based on this expression with the proper boundary conditions applied to each specific geometry. As was pointed out in Section 7.2, unlike with liquids, gas bearing behavior depends on the ambient pressure; thus load capacity, for example, rises with a rise in pa . A new dimensionless parameter now makes an appearance which governs gas bearing behavior given by: =

6µω pa


2 R C

(7.42)

called the Bearing number. Thus, along with geometry and such variables as (L /D ) ratio, load, speed, etc., the value of p, or, in dimensionless form, , constitutes now an additional input. 7.6.1.1 Full Circular Bearings These bearings are the most commonly used if for no other reason than that they are easy to manufacture requiring no grooves or holes for lubricant supply. In obtaining a solution it is only required that at the sides of the bearing the hydrodynamic pressures fall to ambient pressure p, in most cases the atmosphere. Figure 7.38 and Figure 7.39 give the load capacities and frictional losses for full (360◦ ) gas journal bearings for a range of (L /D ratios from 1/2 to 2) and for the entire spectrum of possible 7 values. It can be shown that when 7 6 0, the gas bearing solution approaches that of a liquid; thus the solutions in the figures comprise cases from liquid lubricants to gases of very high compressibility. The reason that the (L /D ) ratios range as high as 2 is to compensate for the inherently low load capacity for gas bearings. 7.6.1.2 Noncircular Geometries As was pointed out in Sections 7.4.2.2 and 7.4.2.3, elliptical and 3-lobe geometries are often resorted to because of their higher stability characteristics. This is particularly desirable with gas bearings which tend to become unstable at high speeds. The solutions given in Sections 7.4.2.2 and 7.4.2.3 for these bearings are for centrally loaded cases, that is when the load vector passes midway through the bottom lobe. However, this is not the optimum mode of loading; better results can be obtained when the bearing is so positioned in the housing that the load is made to pass through the bottom lobe at an angle φL , see Figure 7.40, called the load angle. 7.6.1.2.1 Load Capacity There is a rather large number of independent parameters when dealing with noncircular gas bearings. Assuming even, as was done here, that the space or slots between the individual lobes occupy a negligible portion of the arc, that is, assigning to the elliptical bearing two arcs of 180◦ span each, and to the 3-lobe bearing, three symmetrical (they do not have to be equal) arcs of 120◦ each, we are still left with five independent parameters, namely p = p[(L/D), m, β , αβ , ]

© 2006 by Taylor & Francis Group, LLC

7-52

Handbook of Lubrication and Tribology 50 L/D = 1 40

40 L/D = 1/2

30

e = 0.8

WG

WG

30

20

20

0.7

e = 0.8 0.6

10

0

10

0.4 0.2

0.1

1.0

0.6

0 0 0.2

0.6 Λ

0.2 0

0 0.2

0.6 0.5 0.4 0.3 0.2 0.6

1.0

Λ

1/Λ

50

0.2 0

1/Λ

50 L/D = 1 1/2

L/D = 2 e = 0.8

e = 0.8

40

30

40

10

0.5 0.4 0.3 0.2 0 0.2

0.6 Λ

1.0

WG

WG

0.6

0.7

30

0.7

20

0

0.6

0.6 20

0.5 0.4 0.3 0.2

10

0.6

0.2 0

0

0 0.2

0.6 Λ

1/Λ

1.0

0.6

0.2 0

1/Λ

WG = (6pw/paLRΛ); Λ = (6mv/pa) (R/C)2

FIGURE 7.38 Load capacity of full (360◦ ) gas journal bearing. (Taken from Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.)

A solution for any set of these five parameters will yield the load capacity in the form of the Sommerfeld number S, and the line of action of the resultant force, or load angle φL with respect to the geometry of the bearing. Tables 7.14 to 7.18 give the results with regard to load capacity and some of the other bearing performance characteristics. It should be kept in mind that since load capacity means the relation between Sommerfeld number and minimum film thickness, this hmin is provided not by the bearing eccentricity which is unrelated to the surface curvature, but by the value of one of the lobes, that is, hmin = C(1 − )

(7.43)

In order to obtain this hmin we must search from among the two or three lobes the maximum lobe eccentricity ratio. The values of  listed in the tables and figures are these maximum eccentricity ratios. These most often occur in the bottom lobe. In the few cases when the maximum  occurs in Lobe No. 2,

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Principles of Gas Turbine Bearing Lubrication and Design

7-53

2.60 2.40 e = 0.8 2.20 2.00 1.80 e = 0.6 FC pmv R2L

1.60 1.40 e = 0.4

1.20 1.00

e = 0.2 0.80 0.60

L/D = 2

0.40

L/D = 1

0.20

L/D = 1/2

0 0

0.2

0.6

1.0

0.6

0.2

Λ

0

1/Λ

FIGURE 7.39 Friction in full (360◦ ) gas journal bearing. (Taken from Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961.)

a1 a1 2

a2

a2

2

3 e1

R

e

B

R d

U

e2 a3

e e8 1 e3 e2

–a3

U

1

1

fL y

aB

Y fL

FIGURE 7.40 Nomenclature for noncircular gas bearings. (a) The elliptical bearing. (b) 3-lobe bearing.

this can be identified in the Tables from the fact that for the elliptical bearing this would require αβ > 90◦ ; and for the 3-lobe bearing αβ > 60◦ . While in a circular ungrooved bearing the direction of load application is immaterial. This is not so in the case of noncircular designs. The most common mode of bearing operation is with the load vector parallel to the vertical line of symmetry. This is the natural way of mounting the bearing and it

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7-54

Handbook of Lubrication and Tribology Table 7.14

Centrally Located Elliptical Gas Bearings αB at 

 at 

α at 

B

1/2

1

3

1/2

1

3

1/2

1

3

0.1 0.2 0.3 0.4 0.45

75 75 70 50 30

60 60 50 30 —

30 27 20 — —

0.53 0.58 0.67 0.82 0.92

0.56 0.62 0.73 0.87 —

0.59 0.68 0.79 — —

10 19 25 22 14

9 16 18 13 —

5. 7.5 7.5 — —

S at  B 0.1 0.2 0.3 0.4 0.45

G at 

F =

1/2

1

3

1/2

1

3

1/2

1

3

0.283 0.135 0.0772 0.0341 0.0147

0.250 0.117 0.0625 0.0270 —

0.282 0.128 0.700 — —

0.079 0.0703 0.0455 −0.0204 0.0412

0.127 0.125 0.0620 0.0444 —

0.122 0.101 0.0717 — —

3.28 3.41 3.70 4.54 6.27

3.28 3.40 3.76 4.90 —

3.21 3037 3.86 — —

Note: L /D = 1 m = 1/2 Source: Pinkus, O., Trans. ASME, 80, 1958.

Table 7.15

Optimally Located Elliptical Gas Bearings



βB

B



α

S

φL

G

1/2

0.1 0.2 0.3 0.4 0.45

80 80 80 80 80

0.53 0.57 0.63 0.69 0.73

11 20 28 35 37.5

0.297 0.143 0.0879 0.0581 0.0474

−3 −4 −7 −10 −13

0.0798 0.0735 0.0608 0.0408 0.0279

1

0.085 0.10 0.175 0.20 0.29 0.40 0.475

80 80 75 75 75 80 75

0.52 0.53 0.57 0.575 0.64 0.69 0.77

9 11 17 20 26 35 36

0.362 0.308 0.160 0.149 0.0888 0.0619 0.0411

−12 −12 −9 −13 −12 −18 −19

0.137 0.131 0.124 0.120 0.0995 0.0753 0.0204

3

0.09 0.11 0.22

80 80 75

0.52 0.53 0.60

10 12 21

0.449 0.336 0.163

−23 −23 −21

0.198 0.196 0.168

Note: L /D = 1 m = 1/2 Source: Pinkus, O., Trans. ASME, 80, 1958.

is also useful in that it enables two-directional rotation. However, this is not necessarily the optimum arrangement. Applying the load at various angles to the vertical centerline would yield different values of bearing performance. Somewhere an optimum angle exists for the direction of load application and these are shown in Tables 7.15 and 7.17. In practice it means that depending on the parameters S and , the bearing should be rotated with respect to the load vector anywhere from a few degrees to as much as 25◦ and nearly always in the clockwise direction, in order to obtain the maximum load capacity. Figure 7.41 summarizes graphically some of the data contained in the Tables. Table 7.18 is a practical summary of the various implications contained in the previously discussed results. In practice one is usually confronted with the given requirements of speed, load, ambient conditions, etc. In other words, S and  are fixed. Given these parameters, Table 7.18 shows what eccentricities one can obtain by using a circular or noncircular design.

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design Table 7.16

Centrally Located 3-Lobe Gas Bearings αB at 

B 0.1 0.2 0.3 0.4

7-55

1/2

 at 

1

70 67 63 53

55 55 50 39

α at 

3

1/2

1

3

1/2

1

3

44 42 37 —

0.57(2)a

0.56 0.64 0.73 0.85

0.58 0.66 0.76 —

112(2) 106(2) 99(2) 23

8.4 14.9 18.3 17.2

7.1 11.7 13.5 —

0.64(2) 0.71(2) 0.81

S at 

G at 

F at 

B

1/2

1

3

1/2

1

3

1/2

1

3

0.1 0.2 0.3 0.4

0.293 0.134 0.0734 0.0370

0.301 0.135 0.0734 0.0352

0.437 0.194 0.106 —

0.0358 0.0368 0.0370 0.0314

0.0902 0.0860 0.0773 0.0550

0.0736 0.0675 0.0575 —

3.62 3.86 4.30 —

3.60 3.82 4.25 —

3.55 3.76 4.13 —

Note: L/D = 1 m = 1/2 a (2) indicates that minimum film thickness occurs in right-hand lobe.

Table 7.17 

Optimally Loaded 3-Lobe Gas Bearings

B

αB



α

S

φL

G

1/2

0.1 0.2 0.3 0.4

60 60 60 60

0.56 0.62 0.70 0.78

9 16 22 26

0.284 0.128 0.0720 0.0418

9 7 3 −5

0.0358 0.0364 0.0362 0.0336

1

0.085 0.10 0.175 0.20 0.28 0.30 0.40 0.455

55 55 60 60 60 60 60 60

0.55 0.56 0.61 0.62 0.68 0.70 0.78 0.83

7 8 14.5 16 21 22 26 28

0.385 0.300 0.165 0.140 0.0893 0.0810 0.0477 0.0347

1.5 1 −4 −5 −6.5 −8 −13 −17

0.0580 0.0902 0.0564 0.0862 0.0521 0.0790 0.0425 0.0359

3

0.10 0.225 0.30

50 50 60

0.57 0.67 0.70

8 15 22

0.444 0.171 0.121

−6 −7 −17

0.117 0.0655 0.0578

Source: Pinkus, O., Trans. ASME, 80, 1958.

Table 7.18

Comparison of Load Capacity of Various Gas Bearings



S

m=0 circular

1

0.365 0.162 0.866 0.0381 0.365 0.162 0.0860

0.2 0.4 0.6 0.8 0.25 0.475 0.655

3

Ellipticala

m = 1/2

3-lobea m = 1/2

φt =0

Optimum

φ=0

Optimum

0.54(−170) 0.59(−48) 0.67(−12) 0.815(−2) 0.57(−132) 0.65(−37) 0.75(−14)

0.525(−162) 0.56(−40) 0.635(−6) 0.77(+4) 0.535(−114) 0.59(−24) 0.685(−4)

0.55(−175) 0.615(−54) 0.70(−17) 0.84(−5) 0.595(−38) 0.68(−43) 0.80(−22)

0.545(−172) 0.605(−51) 0.685(−14) 0.815(−2) 0.57(−128) 0.645(−36) 0.755(−15)

Note: (L/D) = 1, Values of β for given  and S a Numbers in parentheses refer to percentage reduction in load capacity from a circular design.

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7-56

Handbook of Lubrication and Tribology

L/D = 1 M = 1/2 Circular Ellipitical 3-lobe S 1.0 8 6

Λ=3

4

2 10–1 8 6 4 Λ=1

2 10–2 0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

e

FIGURE 7.41 Load capacity of symmetrically loaded gas bearings. (Taken from Pinkus, O., Trans. ASME, 80, 1958. With permission.)

7.6.1.2.2 Stiffness Characteristics As was done with load capacity and friction, the stiffnesses of the various bearings will be considered at identical Sommerfeld numbers for all the designs. This means that they will be evaluated at different values of S. This is pertinent from a practical viewpoint since the designer wants to know what the stiffness of the bearing will be under the given conditions of load, speed, etc. regardless of where the journal positions itself in the bearing clearance as a consequence of the imposed operating conditions. Since shaft displacement in different directions produce different responses, here the displacement will be considered in the direction of the load vector. Thus here the definition of the spring constant is given by K = (dF /de ), where F is the response force to a displacement along the load line. Table 7.19 and Figure 7.42 give the results. We see immediately the profound improvement in stiffness in the noncircular over the circular design. In the region of low eccentricities where instability usually occurs (low load, high speed) the value of the spring constants for the elliptical and 3-lobe designs are nearly an order of magnitude higher. 7.6.1.3 Special Design Several unorthodox configurations which have in the past been used on high-speed equipment, including automotive gas turbines, are bearings with grooved surfaces and foil bearings. Figure 7.43(a) and (b) show a herringbone grooved journal bearing and two versions of spirally grooved thrust bearings. In both designs, the bearing or runner surface consists of a lattice of grooves and ridges. From a hydrodynamic point of view, the geometry essentially consists of a series of step bearings, though, unlike with conventional steps, these are at an angle to the direction of motion. One of the achievements of such a design is that the fluid is being driven away from the edges of the bearing, minimizing side leakage and raising load capacity. By a proper orientation of the grooves the fluid can be pumped away from either the inner or outer periphery or from both edges. The advantages of these bearings also lie in the fact that, whereas an ordinary geometry

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

7-57

Table 7.19 Values of K = K /2µLN (C /R )3 m = 1/2



S

m=0 circular

1

0.365 0.162 0.0866 0.0381 0.365 0.162

2.8 5.4 12.8 87.2 5.5 6.8

3

Elliptical

3-Lobe

φL = 0

Optimum, φL

φL = 0

26.2 29.4 42.4 96.0 20.0 25.8

10.2 — 26.4 71.6 26.4 31.2

22.4 28.6 37.4 115.6 16.4 46.2

Optimum, φL — 20.0 26.4 152.6 — 40.4

Note: For L /D =  = 1 Source: Pinkus, O., Trans. ASME, 80, 1958.

100 8 6 4 E

6

K (C/R)3 2mLN

10 8

L

K=

2

4 L/D = 1, m = 1/2, Λ = 1 2

C = Circular E = Elliptical; L = 3-Lobe

C

1.0 0.01

2

3

4 5 6 7 8 90.1

2

3

4 5 6 7 8 9 1.0

FIGURE 7.42 Spring constants for symmetrically loaded gas bearings. (Taken from Pinkus, O., Trans. ASME, 80, 1958. With permission.)

has poor stability characteristics for a concentric shaft position ( = 0), the herringbone bearing is superior to a conventional bearing at low eccentricities. These designs can, of course, be used also with liquid lubricants. 7.6.1.4 Hydrostatic Bearings 7.6.1.4.1 Thrust Bearings Hydrostatic gas bearings are subject to an instability called pneumatic hammer. It is therefore necessary that their recess volume be kept to a minimum. Referring to Figure 7.44 this means that r1 and δ are small. The entrance flow area 2πr1 h from the recess into the clearance will then become more restrictive than the orifice area (πd 2 /4) in which case the bearing is said to be inherently compensated. With an inherently compensated bearing the recess pressure po will equal the supply pressure ps and the drop (ps − p1 ) across

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7-58

Handbook of Lubrication and Tribology (a)

(b) Ridge Groove Groove

Shaft

Ridge Pumping from the outside to the center

Ridge Groove Ridge

Groove

Pumping from center outwards

FIGURE 7.43 bearing.

Special bearing designs. (a) Herringbone grooved journal bearing. (b) Spirally grooved thrust

ds

Ps, Supply pressure P0, Recess pressure P1, Inlet pressure Ps, Ambient pressure

d

Ps

Pa

h P0 P1

r2

Velocity boundary layers

r1

Entrance throat

r

X

L

P0 P1

d Velocity profile r1 r2

FIGURE 7.44

Hydrostatic gas thrust bearing.

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Principles of Gas Turbine Bearing Lubrication and Design

7-59

10.0

d ke

5.0

ed ok

ho

h nc

C

U

P

s=

1/2

2.0

)

1.0 P

s=

0.50

P

(

(g–1)ℜT m CDA0PS 2g m9=

10

s

5

=3

.5

0.20

P

s

=2

.5

= Ps

0.10

0

2.

= Ps

0.05

5

1. = Ps

Ps = 1.75

1

1.

25

1.

= Ps

0.02 0.01 0.01

0.05

0.2

0.5

2.0

FIGURE 7.45

100

(r2–r1)/L

1/2

( )

12mA0CD 2gℜT B= g–1 h3Pa

20

5.0

In (r2/r1)/p

Mass flow rate in a hydrostatic thrust bearing.

the bearing film is given by the compressible equation  m = CD A0 ps

2γ (γ − 1)RT




pa ps

2 γ


−

p1 ps

 γ +1  γ

(7.44)

where m is the mass flow of the gas in kg/sec; Ao is the entrance throat area 2π r1 h in m 2 ; T is the gas temperature in ◦ K; and γ is the ratio of specific heats. To avoid pneumatic hammer, gas thrust bearings must be designed with inherent Compensation, for a design with an (r1 /r2 ) ratio of 0.1. Figure 7.45 and Figure 7.46 provide appropriate design charts, given in terms of a parameter B defined as 12µAo CD B= h 3 pa




2γ RT γ −1

1 2

ln(r1 /r2 ) π

 (7.45)

The chart in Figure 7.45 presents the dimensionless flow, m  , for various values of p¯ = (ps /pa ) while Figure 7.46 gives load capacity. Chart 5 to 10 offers values of bearing stiffness K . To use the charts for design purposes one first used the known value of ps to determine B for the maximum obtainable stiffness from Figure 7.47. From Figure 7.46 one then determines the r2 that will carry the required load W. The value of (h/r2 ) is set within the limits of 0.5 × 10−3 < (h/r2 ) < 2 × 10−3 . With r2 and h established, the actual dimensional values of stiffness k and flow rates can be calculated from Figures 7.45 and Figure 7.47. Refinements are possible by a few more iterations.

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7-60

Handbook of Lubrication and Tribology

ke d Un ch ok ed

10.0

W9R =W/(pr 22ps)

Ps = 10

Ch o

5.0

2.0

Ps = 5

1.0

Ps = 3.5 Ps = 2.5

0.50

Ps = 2.0 Ps =1.75

0.20

Ps =1.5 0.10 Ps =1.25 0.05 Ps =1.1 0.02 0.01 0.01 0.02

0.05 0.1

0.2

0.5

B=

FIGURE 7.46

1.0

2.0

5.0

20

10

50

100

12mA0CD In (r2/r1) 2gℜT g-1 ph3Ps

√

Load capacity of hydrostatic gas thrust bearing.

Optimum orifice compensation

Un

Ch

ch

ok ed

r1/r2 = 0.1

P

s=

2.0

10

Optimum inherent compensation 1.0

P

s=

5

0.50 .5 =3 Ps

.5 =2 .0 Ps = 2 .75 P s = 1 .5 Ps =1 Ps

0.20 0.10 0.05

5 .2

=1 Ps

Stiffness factor, K9= – (h dW/dh)/(pr

2 2 ps)

5.0

ok ed

10.0

= Ps

0.02

1

1.

0.01 0.01 0.02

0.05 0.1

0.2 B=

FIGURE 7.47

0.5

1.0

2.0

5.0

10

20

12mA0CD In (r2/r1) 2gℜT g –1 ph3Pa

√

Stiffness of orifice and inherently compensated hydrostatic thrust bearing.

© 2006 by Taylor & Francis Group, LLC

50

100

Principles of Gas Turbine Bearing Lubrication and Design (a)

7-61

(b)

FIGURE 7.48 Two hydrostatic gas journal bearing geometries. (a) Inherently compensated. (b) Orifice compensated.

LD

10–1

Stiffness K =

2

d C* (1+1+2/3d 2 ) (P –P ) s a

1

10–2

Ps/Pa 20.0 10.1 1.1 1.5

1.0& 2.0

1.5 1.1

10–3 10–3

10–2

10–1

Restrictor coefficient Λsj =

1

101

102

3mnd2 √ ℜT (L/D) 2PsC √1+ d2

FIGURE 7.49 Stiffness of a hydrostatic gas journal bearing. (Taken from Reddickoff, J.M. and Vohr, J.H.,“Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969.)

7.6.1.4.2 Journal Bearings Typical configurations for journal bearings are shown in Figure 7.48(a) and (b). The former is an inherently compensated design, the latter has an orifice restrictor. Here, too, the recess must be small, or in the order of 10% of an incompressible fluid pocket. Capillary restrictors are not used because they cause pneumatic hammer. The analysis of these bearings is very complex and possible only with numerical methods. A typical set of performance curves for an L /D = 1/2 is shown in Figure 7.49 and Figure 7.50. The first shows stiffness as a function of a restrictor coefficient s for various ratios of (ps /pa ). This s is similar to the B parameter in thrust bearings for it represents the ratio of fluid film resistance to the resistance of the restrictor. The parameter δ appearing in the coordinates is defined as

δ = (d 2 /4hr1 )

© 2006 by Taylor & Francis Group, LLC

(7.46)

7-62

Handbook of Lubrication and Tribology 1

Flow, m

6mℜg jm pP 2s C 3

Ps/Pa = 20 1.5 1.25 1.1 10–1

10–2

10–3

20.0 1.5 1.25 1.1

10–3

10–2

10–1

101

1

Restrictor coefficient Λsj =

102

2√ℜT

3mnd (L /D) 2PsC √1 + d2

FIGURE 7.50 Flow in a hydrostatic gas journal bearing. (Taken from Reddickoff, J.M. and Vohr, J.H., “Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969.)

which gives the ratio of the throat area for an orifice restrictor to the throat area represented by the restriction of the bearing film. Its inclusion in the figures permits one to use these charts for both modes of restriction. The K values presented are the center stiffness of the bearing ( = 0). Since the stiffness remains essentially constant up to  = 1/2 load capacity of the bearing can be calculated from W = CK ; or since it is not recommended that the bearings operate at higher eccentricities, the load capacity is given by W = 0.5 CK where K is obtained from Figure 7.49.

7.6.2 Compliant Surface Foil Bearings (CFB) In high-speed, high-temperature applications the CFB’s have the advantage of developing higher load capacities than with fixed geometry gas bearings. Since CFB’s have a complex structure consisting of a spring-like substructure with an overlying flexible surface, there exists a great variety of permutations on any given design. The presentation here will start with two basis models of a journal and a thrust bearing, to be followed with some more elaborate geometries. In all cases the lubricant will be that of air. 7.6.2.1 Foil Journal Bearing The solution of this bearing is based on Equation (7.39), coupled with additional expressions accounting for the elastic behavior of the top and bottom surfaces. The journal bearing is portrayed in Figure 7.51 and its solution is based on the following postulates. • The stiffness of the foil is uniformly distributed around the circumference and is linear with the amount of deflection. • The foil is assumed not to “dag” between bumps but to follow the deflection of the bumps. • In response to the hydrodynamic pressures the deflections are local, that is, they depend only on the force acting directly over a particular point. Under the above conditions the variation in h is due to the eccentricity e and the deflection of the foils. We then have h = C + θ cos(θ − φo ) + K1 (p − pa )

© 2006 by Taylor & Francis Group, LLC

(7.47a)

Principles of Gas Turbine Bearing Lubrication and Design (a)

fL-O

w

7-63

w

f L fL

uE

u ud v

b O O1 C fo

uo

Top foil

(b) Spacer block

Shaft

U

S

h

Bump foil

F

(c)

P t t Io

fF

fF

S

FIGURE 7.51 Configuration of a foil journal bearing. (a) Nomenclature for foil journal bearing. (b) Basic elements of bearing. (c) Configuration of bump foil.

where K1 is a constant reflecting the structural rigidity of the bumps, given by K1 = (p/pa );

α=

2pa s CE


3 lo (l − v 2 ) t

(7.47b)

K1 is then the compliance of the bearing with the quantities, s, lo , and t portrayed in Figure 7.51(c). 7.6.2.1.1 The Nominal Film Thickness In rigid journal bearings, the minimum film thickness is a clear and fixed quantity. It occurs at the line of centers and its value is constant across the axial width of the bearing. Also, generally, the film thickness

© 2006 by Taylor & Francis Group, LLC

7-64

Handbook of Lubrication and Tribology –(L/2)

Z=0

Journal

(L/2) hmin

huN

huominal huo

u = uN

u = uO

FIGURE 7.52

Minimum and nominal film thicknesses.

b = 120°, e = 0.6, uo = 220° (L/D) = Λ = 1, a = 5 U

)

z=

(L/2

uE

z=

0 uS

FIGURE 7.53

Film thicknesses in a 120◦ bearing pad.

anywhere is constant in the z direction (Figure 7.52). Since in our case pressures cause proportional deflections of the bearing surface, the film thickness in the interior of the bearing, where pressures are highest, will be larger than at the edges (z = ±L /2); also since the maximum pressures occur near the line of center, the film thickness is the interior of the θ = θo , the film thicknesses are larger than along another angular position θ = θN = where, because the pressures are lower, the film thickness, on the average, is smaller than at the line of centers. Figure 7.53 shows a three-dimensional film thickness plot for a 120◦ pad in which, while film thickness at the edge (z = 1/2) is small over most of the pad area, the surface has been deflected into much larger values of h. For these reasons a nominal film thickness hN (Figure 7.54) will be defined as the minimum film thickness that occurs along the bearing centerline, that is, at z = 0 at various values of α. While hmin for the rigid case occurs at θ = 180◦ , with increasing values of α the value of this hmin , or our hN , shifts downstream and increases in value; at α = 5, it is twice the value of the rigid case and has shifted downstream by nearly 100◦ . This should be kept in mind later on, when load capacity, that is, the W — hN

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Principles of Gas Turbine Bearing Lubrication and Design

7-65

1.4 1.2

10

hN

5 1.0

hN 1

0.8 h

hN

a=0

0.6 hN 0.4 0.2 0 40

L/D = 1 u0 = 180° z=0 80

Λ = 1.0 b = 180° e = 0.6

120 160 200 240 280 320 360 u, degrees

FIGURE 7.54 Location of nominal film thickness. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.)

relation is plotted; an increase in load while increasing  may also produce an increase in the nominal film thickness. 7.6.2.1.2 Active and Effective Bearing Arc Compliant foil bearings suffer a penalty in their ability to generate hydrodynamic pressures whenever the pad arc commences in a diverging region. The effect can be seen in Table 7.20 which show that by shifting in a 360◦ bearing the line of centers from 180 to 270◦ , there was a loss in load capacity of nearly 30% as well as a reduction in hN . In designing a foil bearing, if the eccentricity if fixed for the particular application, it is best to start the bearing at θ1 = φ (for a vertical load); if the eccentricities are liable to vary, some compromise value of θ1 = 0 can be chosen. 7.6.2.1.3 Performance Characteristics There are six geometric, structural, and operational parameters relevant to a foil journal bearing. These are β, α, (L /D ), , φL , and number of pads. There is also the eccentricity ratio and the attitude angle φ, the latter tied to the load angle φL . A set of standard conditions consisting of (L/D) =  = α = 1;

 = 0.6

is used, and any parametric variation commences from this set of reference values. 7.6.2.1.4 The Full Bearing Table 7.21 gives a detailed listing of the performance of a vertically loaded (φL = 0) full 360◦ bearing as a function of (L/D), α, and . Note should be taken of the fact that the start of the bearing, that is θs is so chosen as to avoid idle (p = ps ) regions at the upstream portion of the bearing. In effect, this requires that θs = φ. The case of nonvertically loaded foil bearings φL = 0, is given in Table 7.20. Some of the noteworthy points emerging from these tabulations are: • Effect of α: While in terms of , there is a drastic drop in load capacity with a more compliant bearing, in terms of hN there is actually an increase in load capacity.

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7-66

Handbook of Lubrication and Tribology Table 7.20 θ (deg)

Effect of Load Angle in 360◦ Bearings

φL (deg)

φ (deg)

θ2 (deg)

Pmax

hN

W × 102

T × 10

— 163.1 141.6 129.8 107.2 33.1 −11.9 −41.9 −72.9

— 4 11.6 20.2 27.2 33.1 28.1 23.1 17.1

— 32 74 95 142 240 272 300 325

0 1.018 1.087 1.133 1.201 1.253 1.245 1.228 1.201

0.40 0.41 0.48 0.52 0.58 0.62 0.62 0.61 0.58

0 1.04 11.5 21.4 40.0 56.8 54.9 50.0 40.1

— 43.9 38.1 33.9 27.2 24.6 23.7 23.7 25.5

0 20 45 60 90 180 220 245 270

Note:  = 0.6; (l /d ) =  = α = 1; θ1 = 0, θE = 360◦ Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.

At large values of α, α > 10, the load the bearing can support is low, due to the fact that the flexible foil deflects sufficiently to maintain high film thicknesses even at large eccentricities. Thus from a design standpoint, it may be advisable to use high-compliance bearings at low loads; high loads, however, can be supported only with bearings of low values of α. In highly compliant bearings (particularly at high L /D ratios) an increase in eccentricity may produce an increase in hN , a phenomenon opposite to rigid bearings where hmin is the inverse of . • Effect of : The performance of a foil bearing as a function of  conforms to the familiar pattern ˜ with an increase in , the load capacity, of compressible lubrication. After an initial rise in W ˜ both in terms of an increase in W as well as a rise in hN , tends to flatten off and approach an asymptotic value. The torque, however, rises almost as a linear function of the increase in . The more compliant bearing shows lower power losses due to the prevailing higher film thickness. 7.6.2.1.5 The Multipad Bearing The 3-pad design consists of three 120◦ arcs and the 5-pad design has five 72◦ arcs. In each case the vertical line of symmetry dissects the bottom pad, so that φL = 0 represents a load passing through the midpoint of the bottom pad (Table 7.22). Tables 7.23 and 7.24 give a spectrum of solutions for the performance of the 3-pad bearing and these results show the following: • Variation with load angle: Because of the cyclic nature of this bearing (symmetry for each 120◦ ) ˜ or T with a shift in load angle. In particular, there is no there is much less variation in either W acute loss of load capacity when the line of centers passes between pads. The optimum load angle for α = 1 is φL = −10◦ and for α = 5 it is φL = −14◦ . The improvement in load capacity over that of central loading (φL = 0) is of the order of 10 to 15%. • Variation with number of pads: Figure 7.55 shows the variation of 1-, 2-, and 3-pad bearings as a function of load angle. The plot shows clearly a drop in load capacity with the number of pads, that is, with a drop in extent of bearing arc β. As seen, the optimum for the 360◦ bearing occurs at φL = 0, at which point the torque also reaches it minimum value. The 3-pad bearing, as said previously, reaches an optimum at φL = −10◦ ; whereas the 5-pad bearing reaches an optimum at φL = −15◦ . 7.6.2.1.6 Stiffness Table 7.24 gives the values of the four spring coefficients for two values of compliance, the limiting case of α = 0 and α = 1. The α = 0 case differs from a rigid gas bearing in that the subambient pressures are eliminated from the pressure profile. A comparative evaluation of the stability characteristics of the 1- and 3-pad bearings is, of course, best done in a study of a rotordynamic system, particularly when the

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design Table 7.21

7-67

Performance of a 360◦ Foil Journal Bearing φ = θs

θ2 − γ

P max

hN

W × 102

T × 10

L/D = 0.5 0.3 0 1 5 10 20 0.6 0 1 5 10 20 0.9 0 1 5 10 20

63.5 59.0 48.5 60.0 34.0 40.0 36.0 32.0 30.0 27.0 12.0 19.0 21.0 21.0 21.0

81.5 87.0 100.5 114.0 114.0 62.0 76.0 97.0 105.0 114.0 52.0 71.0 91.0 99.0 108.0

1.046 1.043 1.037 1.025 1.017 1.25 1.144 1.073 1.05 1.033 3.73 1.33 1.12 1.077 1.048

0.70 0.72 0.80 0.85 0.94 0.40 0.51 0.68 0.795 0.90 0.10 0.41 0.64 0.76 0.91

4.4 4.2 3.2 2.9 2.1 17.9 13.7 8.3 6.1 4.2 157.3 34.7 14.8 9.8 6.3

9.34 9.15 8.53 8.09 7.54 15.85 13.93 8.19 7.33 6.54 26.1 13.8 9.8 8.5 7.3

L/D = 1.0 0.3 0 1 5 10 20 0.6 0 1 5 10 20 0.9 0 1 5 10 20

37.0 49.0 36.0 28.0 20.0 36.0 33.0 28.5 25.0 20.0 13.0 21.0 23.0 21.5 19.0

97.0 104.0 117.0 120.0 132.0 77.0 95.0 112.0 117.0 120.0 59.0 86.0 94.0 108.5 127.0

1.173 1.107 1.061 1.041 1.025 1539 1.253 1.114 1.074 1.046 4.850 1.434 1.154 1.103 1.063

0.70 0.77 0.94 1.04 1.14 0.40 0.62 0.90 1.055 1.22 0.10 0.52 0.86 1.05 1.26

27.9 23.7 14.8 10.3 6.37 94.9 56.8 28.8 19.4 12.2 504.5 102.8 42.9 27.8 17.2

22.7 21.2 18.6 17.5 16.5 31.1 24.6 19.1 16.9 15.0 58.2 28.1 19.5 16.7 14.4

L/D = 1.5 0.3 0 1 5 10 0.6 0 1 5 10 0.9 0 1 5

52.0 43.0 29.0 21.0 35.0 32.0 26.0 22.0 14.0 23.0 23.0

103.0 113.0 119.0 141.0 88.0 104.0 120.0 137.0 68.0 95.0 112.0

1.218 1.152 1.076 1.048 1.731 1.311 1.135 1.084 5.300 1.485 1.184

0.70 0.82 1.03 1.13 0.40 0.68 1.00 1.18 0.10 0.56 0.96

70.0 53.2 28.5 18.3 208.9 112.0 52.0 34.1 298.9 179.7 74.2

33.4 30.4 26.4 24.8 45.6 34.3 26.3 23.1 85.1 39.2 26.7



α

Note:  = 1; φL = C Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.

cross coupling components vary not only in magnitude but also in sign. However, the following items can be deduced from the tabulated  data: ˜ the Kyy ’s are about the same for both designs, whereas the Kxx ’s are lower • When plotted against W for the 3-pad configuration.

© 2006 by Taylor & Francis Group, LLC

7-68

Handbook of Lubrication and Tribology Table 7.22

P Performance of a 3-Pad Bearing (L/D) =  = 1, B = 120◦ θs

θL

φ

P max

W × 102

T × 10

40 210 217 220 225 245 273

−140.0 7.2 −2.3 −2.3 −5.2 −10.0 −17.4

79.2 37.2 27.0 27.7 29.8 55.0 77.6

1.073 1.072 1.075 1.079 1.082 1.088 1.077

12.1 12.7 13.7 14.1 14.6 15.0 12.6

21.4 21.8 21.9 22.0 22.0 21.5 21.4

29 180 200 2202 245 270 15

−145.0 44.1 −2.6 −10.6 −17.0 −25.5 −145.0

69.2 44.1 30.6 29.4 48.0 64.5 50.9

1.188 1.133 1.197 1.215 1.284 1.185 1.497

24.0 25.2 37.2 29.8 34.9 28.7 59.3

27.8 18.6 27.7 28.8 29.0 27.4 62.9

194 210 230 245 260

0.0 −10.8 −14.1 −14.7 −26.2

16.3 19.2 333.9 50.5 53.8

1.340 1.375 1.572 1.412 1.463

69.5 76.9 74.3 52.4 55.3

34.7 45.7 71.8 77.3 56.6

38 210 214 220 225 245 275

143.0 5.1 0.0 −5.3 −8.6 −13.3 −11.1

74.9 35.1 34.3 34.7 36.4 51.7 73.9

1.049 1.038 1.040 1.043 1.045 1.053 1.050

0.6

12 180 205 220 245 270

145.0 55.0 3.6 −9.7 −16.3 −23.1

67.4 55.5 28.6 30.3 48.74 66.9

1.104 1.048 1.077 1.103 1.121 1.106

15.7 12.6 16.1 18.3 18.4 15.9

25.4 16.5 25.3 26.7 27.2 25.4

0.9

25 200 203 210 230 245 260

164.0 3.5 0.0 −4.3 −16.3 −18.9 −21.5

61.0 23.5 23.0 23.7 33.7 46.1 58.5

1.178 1.122 1.119 1.158 1.198 1.202 1.186

24.4 25.3 26.3 28.1 29.7 18.2 25.2

46.5 34.4 36.2 41.6 67.1 72.1 53.7

 α=1 0.3

0.6

0.9

α=5 0.3

Source:

0.01 7.52 7.87 8.31 8.60 9.07 8.18

20.7 21.1 21.1 21.2 21.2 20.8 20.7

Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.

• With the more compliant case, the K ’s tend to level off with a rise in eccentricity, the values of the coefficients approaching the structural stiffness of the system. In general, the advantage of the compliant bearings in the area of stability lies in that levels of stiffness can be selected by the designer via a proper combination of structural and hydrodynamic stiffnesses. Thus, instead of making his inertias and supports suit the inherent stiffnesses of purely hydrodynamic bearings, the designer may try to tailer and adjust bearing stiffness to the demands of high rotordynamic system.

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design Table 7.23

7-69

Mode of Load of 3-Pad Bearing

α



φ

W

φL

φ

W

1 1 1 5 5 5

0.3 0.6 0.9 0.3 0.6 0.9

37.5 28.5 16.0 35.0 30.0 23.5

13.8 36.0 68.0 7.8 17.0 26.2

−10 −10 −10 −14 −14 −14

55.0 29.0 18.5 53.0 41.0 30.0

15.0 39.8 78.0 9.0 18.6 29.8

Note: (L/D) = λ = 1; β = 120◦ each Central loading φL = 0; Optimum loading Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of GasLubricated Foil Journal Bearings,”ASME Paper 82-LUB-40, 1982.

Table 7.24 Values of Spring Coefficients 

α

 = 360◦ 0.6 0 0.75 0 0.9 0 0.6 1 0.75 1 0.9 1

φ

W

Kxx

Kxy

Kyx

35.7 24.1 12.8 32.1 26.3 21.4

0.951 1.894 5.055 0.568 0.7833 1.028

1.920 3.416 7.202 1.129 1.231 1.268

−0.125 −1.166 −6.024 0.174 0.0254 −0.098

−2.345 −3.989 10.151 −0.693 −0.686 −0.627

3.237 8.981 44.593 1.130 1.378 1.602

0.635 1.321 3.695 0.359 0.511 0.689

1.123 2.102 4.728 0.5702 0.673 0.759

−0.092 −0.752 −3.344 0.0451 −0.017 −0.057

−2.05 −3.710 −8.768 −0.758 −0.821 −0.855

2.635 7.432 37.103 0.801 1.051 1.274

3-pad-120◦ each 0.6 0 26.0 0.75 0 17.4 0.9 0 8.6 0.6 1 25.5 0.75 1 20.5 0.9 1 16.3

Kw

Note: (L/D) = ; φL = 0 Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.

0.80

L/D = 1 a=1

0.70

Λ=1 e = 0.6 1 Pad —360° 3 Pads —120° 5 Pads —72°

0.60

hN

0.50 0.40 W 0.30

W T=10

0.20 W

0.10 0 180

140

100

60

20

–20

–60

–80

FIGURE 7.55 Performance of multipad bearings. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.)

© 2006 by Taylor & Francis Group, LLC

7-70

Handbook of Lubrication and Tribology (a)

(b)

U h2 h1

U Pressure distribution

Original surface

h2

h1

R2

L R1

bb B u

hN

b

Deflected surface

FIGURE 7.56 The geometry of compliant surface thrust bearing. (a) Nomenclature of the thrust bearing. (b) The elastohydrodynamics of a complaint foil bearing. Table 7.25

Effect of β on Bearing Performance

h1

α◦

◦

β

n

W × 102

T × 102

p max

W TOT × 102

2.5

4

1.2

3.0

3-1/3

1.0

4.0

2.5

0.75

30 45 60 75 30 45 60 90 30 45 60 90

12 8 6 5 12 8 6 4 12 8 6 4

0.259 0.430 0.573 0.686 0.257 0.435 0.588 0.808 0.238 0.415 0.574 0.814

4.19 6.09 8.02 9.98 4.13 5.97 7.81 11.58 4.01 5.82 7.56 11.08

1.0288 1.0320 1.0329 1.0331 1.0293 1.0334 1.0349 1.0349 1.0285 1.0336 0.0359 1.0373

3.12 3.44 3.44 3.30 3.08 3.48 3.53 3.23 2.86 3.32 3.44 3.26

Note: (L/R2 ) = 0.5; b = 0.5 Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982.

7.6.2.2 Thrust Bearings Figure 7.56 shows the configuration of the thrust bearing considered next, which resembles that of a conventional tapered land design. All the postulates stated in connection with the journal bearing apply here as well, except, of course, that the film thickness is different. This is now given by h = h2 + g (r, θ ) + C(p − pa )

(7.48)

where (h1 − h2 )[1 − θ/bβ} g (r, θ) = 0

for 0 ≤ θ ≤ bβ

for bβ ≤ θ ≤ β

7.6.2.2.1 Geometric Optimization As shown in Table 7.25 the maximum pressure changes vary little with an increase or decrease in β or the value of h1 . The table also shows that when the total number of pads possible for a given value of β is accounted for in the calculation of the total load capacity, there is little difference in the choice of a

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Principles of Gas Turbine Bearing Lubrication and Design

7

(L/R2) = 0.5 Λ = 3.0b = 0.8 a = 1.0 h1 = 2.0 W =10

(b) 10

3.5

9

W TOTAL

3

8

6

2.5

W Pad 102

7 5

W

6

WToT =102

(a)

7-71

2

5 4

4

3 2

3

1 20° 30° 40° 50° 60° 70° 80° 90°

L/R2 = 0.5 b = 0.5 Λ = 1.33 a = 4.0 h1 = 2.5 hu =1.31= constant b.Degrees 30 40

50

60

70

80

90

FIGURE 7.57 Effect of β on pad and total load. (a) Effect of β in stiff bearings. (b) Effect of β in soft bearings. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)

particular arc length. The effect is shown in Figure 7.57 for both a relatively stiff and soft bearing. In both cases the nominal film thickness does not vary with β. The total load capacity in a full 2π thrust bearing shows a maximum somewhere between 45 and 50◦ . The effect of the proportion of ramp is shown in Figure 7.58. For a relatively stiff bearing the optimum is about 70%, which is close to the case of rigid bearings. For the more compliant case the optimum is 50%. For high values of , the optimum value of b recedes to values closer to 40%. It can be seen that near the highest values of W the nominal film thickness is also at its highest, thus reinforcing a general load capacity optimum for compliant bearings at about b = 0.5. ˜ changes little with a The effect of varying h1 is given in Table 7.26. It shows that while the value of W ˜ variation of h1 , the nominal film thickness goes up appreciably, doubling in value for a doubling of h1 . Thus, in terms of customary load capacity criteria, the highest values of h˜ 1 seem desirable. To summarize, the optimum geometry for a bearing with the common OD to ID ratio of 2 is β = 45◦ ,

b = 0.5,

h˜ 1 > 10

7.6.2.2.2 Performance Characteristics The performance of a CFB for the optimized parameters of β = 45◦ and b = 0.5 are given in Table 7.27 for a wide range of parameters h1 , α ◦ , and ◦ . In terms of h2 normalization the range of  extends to nearly 1000 and that of α to over 60. Figure 7.60 is a performance plot in terms of the basic variables involved in the bearing. The drop of load capacity with an increase in film thickness and with a decrease in the value of  are trends known from other studies of gas bearings. What is particularly noteworthy in Figure 7.59 is the effect of α on load capacity. While at moderate ’s high values of α yield the highest load capacity at high , the optimum α is some intermediate value, in our case α ◦ = 1. Note should ◦ are all normalized by the geometric ramp height δ; be taken that all quantities, that is, ◦ , α ◦ , and hN and that the Figure 7.59 plots contain implicitly various values of h1 . The relation between h˜ 1 and h˜ N is given in Figure 7.60. This graph can be used to determine various ramp heights for different points of Figure 7.59. The h˜ 1 − h˜ N graphs support the conclusion of the previous section as to the desirability of using high values of h˜ 1 since they yield high nominal film thicknesses. It also shows that the higher α ◦ and ◦ are, the higher the film thickness.

© 2006 by Taylor & Francis Group, LLC

7-72

Handbook of Lubrication and Tribology 10 (L/R2) = 0.5 b = 45° h1 = 2.0 a* = 1.0 Λ* = 3.0

9 8 7

T W × 103 T × 102

6

5 W 4 3

1.1

2 hN

hN 1.0

1 0

0.2

0.4

0.6

0.8

1.0

b

FIGURE 7.58 Effect of extent of ramp of performance of CS thrust bearing. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.) Table 7.26

Effect of h 1

h1

W × 102

T × 10

hN

10 15 17 19 20

8.169 8.626 8.730 8.811 8.845

44.5 72.2 83.3 94.5 100

5.106 7.675 8.730 9.730 10.265

K × 103 114 116 116 117 122

Note: β = 45◦ , L /R2 = 0.5, b = 0.5, α a = 1.0, a = 10. Source: Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.

Finally, Figure 7.61 gives a plot of the spring constant for the bearing. The stiffness of the bearing is, of course, a function of both the structural stiffness as represented by KB and of the hydrodynamic film stiffness. Since they are in parallel, high loads would tend to flatten the values of K for the softer bearings, leaving essentially the structural stiffness KB as the dominating spring constant. 7.6.2.3 Advanced CFB Designs The construction of CFBs lends itself to a number of modifications that can enhance a particular performance characteristic in accordance with operational requirements. Due to their analytical complexity and space limitations they cannot be gone into to any extent but a mere listing of some of them will give an idea of the range of possibilities latent in this group of bearings.

© 2006 by Taylor & Francis Group, LLC

Compliant Foil Thrust Bearing Performance α=0 W × 102

a hN

h1 = 2 0.1 1.0 10.0 20.0 40.0

0.014 0.144 0.632 3.33 6.246

h1 = 5 0.3 1.0 10.0 20.0 40.0

α=1

T × 10

K × 103

W × 102

a hN

1.0 1.0 1.0 1.0 1.0

0.024 0.241 2.40 4.79 9.50

0.314 3.24 38.6 76.4 125.0

0.104 0.145 1.52 2.736 4.238

0.192 2.343 22.2 34.66 46.4

0.25 0.25 0.25 0.25 0.25

0.35 3.50 33.5 65.2 127.0

12.2 171.0 1230.0 1730.0 2060.0

h 1 = 10 0.1 1.0 10.0 20.0

0.66 8.818 60.38 86.84

0.111 0.111 0.111 0.111

1.63 16.1 150.0 292.0

h 1 = 20 0.1 1.0 10.0 20.0

2.084 22.35 134.8 185.5

0.0526 0.0526 0.0526 0.0526

6.60 64.6 605.0 1179.0

a

α=4

α = 20

T × 10

K × 103

W × 102

a hN

T × 10

K × 103

1.001 1.011 1.114 1.197 1.276

0.024 0.24 2.27 4.34 8.21

0.314 3.20 31.5 45.5 51.5

0.014 0.147 1.214 1.828 2.406

1.004 1.046 1.326 1.461 1.574

0.024 0.23 2.04 3.78 7.05

0.32 3.35 17.5 18.8 16.4

0.204 1.833 6.633 8.135 9.231

0.268 0.378 0.635 0.71 0.758

0.341 2.72 17.3 30.8 57.0

14.0 77.7 104.0 95.7 78.0

0.206 1.164 2.697 3.362 3.612

0.316 0.538 0.905 0.978 1.015

0.31 2.07 12.9 28.6 44.9

91 1,120 6,120 8,240

0.661 2.170 8.109 9.457

0.161 0.307 0.566 0.630

1.30 7.81 44.5 79.5

61.1 113.0 114.0 95.6

0.571 2.309 5.345 6.020

0.192 0.38 0.695 0.785

575 4,720 26,600 34,200

1.103 3.876 8.845 10.24

0.16 0.282 0.54 0.596

3.54 18.3 100.0 179.0

90.0 128.0 117.0 97.5

0.843 2.705 5.674 6.280

0.162 0.364 0.666 0.722

a

a

W × 102

a hN

T × 10

2.3 29.8 26.1 19.7 14.7

0.152 0.513 0.854

0.452 0.834 1.15

0.24 1.43 10.2

5.24 6.92 3.64

9.915

1.18

38.8

1.16

1.15 6.50 37.5 68.0

40.5 64.8 56.7 45.0

0.464 1.619 3.323 3.627

0.234 0.485 1.825 0.88

0.98 5.38 32.3 59.8

25.0 35.6 25.0 18.5

2.91 14.8 83.6 152.0

52.2 70.6 56.2 44.0

0.623 1.83 3.467 3.733

0.209 0.404 0.790 0.835

2.39 12.1 71.7 113.0

29.7 37.8 24.3 17.7

a

K × 103

Principles of Gas Turbine Bearing Lubrication and Design

Table 7.27

Note: β = 45◦ , b = 0.5, (L/R2 ) = 0.5

7-73

© 2006 by Taylor & Francis Group, LLC

7-74

Handbook of Lubrication and Tribology 100

b = 45* (L/R2) = 0.5 b = 0.5 a* = 0 a* = 1 a* = 4

10

— W × 102

Λ* = 20 1

Λ* = 1

0.1

Λ* = 0.1

0.01 0

0.2

0.4

0.6

0.8

1.0

1.2

1.4

h*N

FIGURE 7.59 Load capacity of CS thrust bearing. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)

15 b = 45* b =0.5

13

(L/R2) = 0.5

20 10

a* = 1 a* = 4

11

20

Nos. refer to Λ*

10

hN

9 1 7 1 5 0.1

3 a = 0 All Λ*

1 1

3

5

7

9

11

13

15

17

19

21

h1

FIGURE 7.60 Relation between h1 and hN . (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

7-75

10 b = 45* (L/R2) = 0.5 b = 0.5 a* = 0 a* = 1 a* = 4

5 2 1 5 2

Λ* = 20

10–1

0.1

5 Λ* =

2 –2

=1

10

Λ*

5 2 10–3 5 2 10–4

— W –4

10

10–3

10–2

10–1

1

FIGURE 7.61 Stiffness of CFB. (Taken from Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Thrust Bearings,” ASME Paper 82-LUB-39, 1982.)

7.6.2.3.1 CFBs With Variable Direction Stiffness The bump foil design can be manipulated to provide a wide range of desirable dynamic properties. One such arrangement to vary the stiffness in both radial and circumferential directions is shown in Figure 7.62. The stiffness gradient permits the formation of a variable hydrodynamic wedge in accordance with variation in load or speed. As speed increases the particular arrangement can be made to increase film convergence which enhances stability precisely when it is needed. In one such application an advanced air-lubricated journal bearing reached speeds of 132,000 rpm carrying a unit load of close to 100 psi. Much of the above applies also to thrust bearings. A proper thrust bearing geometry is one that has a taper in the circumferential direction followed by a flat portion. A CFB can be constructed with a foil possessing stiffening elements at the trailing edge, as shown in Figure 7.62(b). The stiffening elements placed between the top and bump foils provide a variable stiffness gradient from the leading to trailing edge yielding the desired converging shape. 7.6.2.3.2 CFBs with Controlled Coulomb Damping A foil bearing can be constructed to improve internal damping and thus enhance its stability characteristics. This can be done by affecting the Coulomb damping due to the relative motion between the top and bump foil surfaces, as well as between the bump foil and the housing (see Figure 7.63). This relative motion occurs, of course, when the bearing is loaded and the foils are radially deflected. To improve the friction characteristics of this relative motion the rubbing surfaces are sputter coated with copper, silver, or some other high friction material. Sometimes the surfaces of the journal and mating top foil are coated with dry lubricants to minimize friction on start-up and shut down. To further enhance stability the bump foil

© 2006 by Taylor & Francis Group, LLC

7-76

Handbook of Lubrication and Tribology (a)

W

W + DW

(b)

Multilayered top foil

Runner

Variable pitch supports

FIGURE 7.62 Variable support stiffness in compliant bearings. (a) Variable radial stiffness. (b) Variable longitudinal stiffness.

Top foil

Dry film coating

Copper coating

Relative motion under detormation

Bump foil Housing

FIGURE 7.63

Mechanism for Coulomb damping in foil journal bearing.

can be circumferentially split along axial lines to improve alignment and axial compliance. A single pad design of this variety carried unit loadings up to 100 psi. 7.6.2.3.3 CFBs With Cantilevered Leaves A bending dominated foil bearing is the cantilevered leaf type design. A journal bearing of this type is shown in Figure 7.64(a). Here each leaf is preformed with a specific radius to induce a desirable film profile. The thrust bearing variety is shown in Part b. It is essentially a thin plate with springs mounted below it which is separated by a backing plate from a foil assembly plate to which the leaves are attached. The plates are in the form of annular rings and the leaves in the form of annular sectors. The leaves should be thin enough to permit considerable flexibility but still thicker than the hydrodynamic film.

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design (a)

Journal

(b)

7-77 Thrust Foil segment removed

Foil assembly

Foil housing Assembled foil segments

Blacking plate Foil segment removed

Spring assembly Foil

FIGURE 7.64

Foil

Garrett self acting CFB.

7.6.3 Magnetic Bearings To illustrate the potential of magnetic bearings, a specific design example will be followed through. It will provide orientation and a guide for general cases where their use is contemplated, particularly with regard to controlling instability and resonances at high speeds. 7.6.3.1 General Principles Practical MBs are mostly the attractive type. Radial AMBs generally adopt an 8-pole stator configuration as shown in Figure 7.65. Both the stator and journal consist of laminations of ferromagnetic material. The journal is shrunk on a shaft without windings. The use of laminations reduces eddy current, which not only causes power loss, but degrades the dynamic performance of the bearing. The stator poles are separated into four quadrants. In each quadrant, the electromagnetic windings are wound in such a way that the magnetic flux circulates mainly inside the quadrants so that each quadrant of poles can be controlled independently. Magnetic force is proportional to the ratio current to air-gap squared (Figure 7.66). To support a load in a controlled axis (Figure 7.67), unequal steady state or bias currents are induced in opposite pairs of poles, such that W = f (I12 − I32 )/C 2

(7.49)

where I1 > I3 f = a magnetic pole constant for a given number of windings. The bias current produces an I 2 R loss, which is a major power loss in an AMB. However, the total resistance in the current path is not large; the AMB loss is, in general, insignificant compared to fluid film bearings. The journal floating in the magnetic field due to its bias currents alone is stable. Linearized

© 2006 by Taylor & Francis Group, LLC

7-78

Handbook of Lubrication and Tribology A quadrant Current Windings

Flux path

Air gap Shaft

Laminated journal

Laminated stator

FIGURE 7.65 An 8-Pole configuration of an active magnetic bearing.

Mafnetic force

Air gap decreases

Designed air gap

Air gap increases

Linearized feedback current operating point

Bias current

FIGURE 7.66 Nonlinearity of magnetic force. (Taken from Allaire, P.E., Li, D.F., and Choy, K.C., J. Lubr. Technol., Trans. ASME, July 1980. With permission.)

feedback control is achieved by making the air gap large relative to the journal’s vibrations. For stable operation the journal motion must be sensed and corrected instantaneously and continuously by superimposing a small control current to each bias current. For example, as shown in Figure 7.67, when the journal moves up by a small displacement, the current in the top quadrant will be reduced to a small amount, i, and the bottom quadrant increased by i. The control currents produce a net downward force, F , which pulls the journal back to the center. From sensing Y to producing F , a series of AMB components are involved, namely sensor, controller, power amplifier, and electromagnets.

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

7-79

Measured journal displacement Bias Ib1 adjustment

Power amplifier No.1 Ib1 – i Stator

Air gap Y Controller X Shaft –1 Journal

Stator Bias Ih3 adjustment

Ib3 + i

Power amplifier No.2

FIGURE 7.67 An independently controlled axis.

7.6.3.2 AMB Components 7.6.3.2.1 Electromagnets For a given maximum load including static and dynamic loads, the AMB physical size is determined by the saturation flux density of the lamination material. For the 8-pole configuration, Fmax = 5.75 × 105 Ap Bs2

(7.50)

where, Fmax = maximum load, N Ap = surface area per pole, m2 Bs = saturation flux density, W/m2 The maximum value of the flux density in the linear range which is about 90% of the actual Bs value should be used in applying Equation 7.50. Choosing the axial length Lp , the circumferential pole width is Ap /Lp . The radial dimensions can be determined from a given shaft diameter at the AMB. For sizing, the following guidelines should be followed: • • • •

The cross-sectional area at any point of the flux path is not less than Ap . Adequate wiring space is provided. The axial length is no greater than the journal OD. As a rule, the air gap should be ten times the expected journal vibration.

The pole surfaces are the most effective areas for heat dissipation by convection. The ampere-turns per pole is fixed for a given ferromagnetic material; the optimal choice of winding turns, Mt is a trade-off of total current and inductance load, L, to the power amplifiers. The latter is proportional to Mt2 ApC which is a crucial parameter causing control delay and bearing instability. More than 8 poles can be designed for the stator, such as 16 or 24 evenly spaced. A large number of poles saves radial space because it localizes flux circulation. The coil pairs can be in series or parallel to a power amplifier with the same trade-off.

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7-80

Handbook of Lubrication and Tribology

7.6.3.2.2 Power Amplifiers Converting a low power control voltage signal to a high power control current and actuating the electromagnets requires power amplifiers. Two types are commercially available, the linear and the pulsewidth-modulation (PWM) type. The linear amplifier applies the control signal to a power transistor in an “active mode.” The transistor continuously regulates the current through the windings from a DC source, Vs , with the current directly proportional to the control signal. The PWM type applies the control signal to generate high voltage pulses at a fixed frequency above audible range. The on-time period of each pulse is proportional to the input signal. The voltage pulse train produces current to the windings. The PWM type is electrically noisy and needs its own filters. The power transistors operate in a “saturation mode” with much less power loss. There are three requirements for the power amplifiers. First, the control current, i, cannot be larger than the bias current. Second, the inductance of the electromagnets causes the control current to diminish and delay above a certain frequency (cut-off frequency). The PWM amplifiers usually apply their own current feedback to increase this frequency. Third, the value of Vs /L, called the current slew rate limit, is the maximum amperes per second that the amplifier can provide. 7.6.3.2.3 Sensors Three displacement sensors prove to be practical, the capacitance probe, inductance, and eddy current probe. Each varies with its advantages and disadvantages, but all relate the small distance between the stationary sensor and the rotating shaft to an output electrical signal in volts. A low-pass filter is usually included in the sensor conditioning device to eliminate high frequency noise, including its own FM carrier. This filter, similar to the power amplifier cut-off characteristics, may cause a significant time delay in the frequency range of interest. A phase-lead circuit implanted in series in the feedback loop can reduce the delay. An inexpensive and reliable sensor is not yet available for measuring the journal velocity. Different analog circuits, such as a differentiator with a low-pass filter, and phase-lead circuits have been used to produce a pseudo velocity from the displacement measurement. An analog surrogate called a Velocity Observer, instead of differentiating displacement, integrates journal force (equivalent to acceleration) to obtain velocity. The output of any pseudo velocity circuit is a combination of displacement and velocity signals. Thus, its feedback not only produces damping, but also contributes to the stiffness. 7.6.3.3 AMB Stiffness and Damping From the previous discussion, a practical single axis control can be represented by the block diagram of Figure 7.68. A radial AMB needs two independently controlled axes like this, while a thrust AMB needs only one. A second-order, low-pass filter was assumed to be part of the sensor though it could have been a fourthorder or some other type of filter. Gp is the sensor sensitivity; for example, 1000 V/in. (40 V/mm). A phase-lead circuit is applied in series here for compensating the time delay caused by the inductance loads to the power amplifiers. One may set the phase-lead parameter “a” to be equal to the amplifier cut-off frequency, ωn . Thus a system “zero” cancels a system “pole” in the 3-plane. This does not improve the current slew rate of the amplifier, but does increase the damping-to-stiffness ratio around ωn . The other phase-lead parameter “b” is set in the range of a ≤ b ≤ 10a. The AMB stiffness and damping of this controlled axis can be calculated by using the equations below with S equal to jω.

© 2006 by Taylor & Francis Group, LLC

−F /Y = K + jωb = −Ki (i/Y ) − Km

(7.51)

i/Y + (TS )(TC )(Tp )(Ta )

(7.52)

Principles of Gas Turbine Bearing Lubrication and Design

7-81 Journal motion at AMB CL

Adjustable PID gains Cd

Y

Gpvc2

Yp

S2+√2vcS + vc2 Measured Sensor journal displacement low-pass filter

vo S + vo

Q

S S + vv

Z

Km

–

Ce

– b S+a × a S+b –

E

Phase-lead circuit

Gava S + va Power amplifier

i

+

+ F

Ki

Electromagnets

Cv

Controller Rotor/AMB force coupling equation F = Kii + KmY AMB state equations i ’ + vai =GavaE E ’ + bE = – (b/a)[CdYp’ +CeQ’ +CvZ ’ +a(CdYp +CeQ +CvZ )] Q’ + voQ = voYp Z ’ + vvZ =Vp Yp’ =Vp Vp’+√2vcVp + vc2Yp =Gpvc2Y

FIGURE 7.68 A single-axis control diagram.

where ω is the excitation frequency and TS = Ga ωn2 /(S 2 −

√

2ωnS + ωn2 )

TC = −Cd − Ca ωo /(S + ωo ) − CV S /(S + ωV ) Tp = (b /a )(S + a )(S + b ) Ta = Ga ωn /(S + ωn ) The equations indicate that both K and B are functions of excitation frequency ω, not rotational speed. Numerical results are plotted in Figure 7.69 using the AMB data. The frequency axis in this plot is normalized with respect to 50 Hz which is the average of two rigid-body critical speeds of a rotor. The amplitude is normalized with respect to (Gp Ca Ki Cd − Km ). At the low-frequency range where the integral control dominates, the plot shows negative damping values. This should not cause alarm, however, since mechanical system resonances seldom exist in that low range. At the high-frequency range, especially where the first two bending criticals exist, negative damping can cause resonances. This is discussed later. 7.6.3.4 Rotor-AMB System Dynamics 7.6.3.4.1 System Design Guidelines Active Magnetic Bearings are generally less stiff than rolling element or hydrodynamic oil-film bearings. Therefore, the first two system criticals have relatively rigid mode shapes, and their vibrations are easily controlled. The third and fourth criticals with bending mode shapes must be given careful consideration in high-speed turbomachines.

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7-82

Handbook of Lubrication and Tribology

Lp = 2.0 in. (50.8 mm) D = 2.5 in. (36.5 mm) C = 0.020 in. (0.5 mm)

50,000

Ap = 1.25 Nt = 100 turns Fmax = 200 lb (890 N) I1 = 3.5 A I3 = 2.0 A

Ki = 80 lb/A (356 N/A) Km = 12,500 lb/in. (2.19 × 106 N/m) Cd = 0.26 Cv = 0.6 Cc = 1.0 Gp = 1000 V/in. Gz = 1 A/V vc = 5000 Hz vo = 1 Hz vv = 500 Hz va = 500 Hz

Dynamic stiffness (lb/in.)

in.2(8.06 × 10–4 mm2) √(K 2 + v2B2) K 0

vB

–50,000 1

10 100 Excitation frequency (Hz)

1,000

FIGURE 7.69 AMB stiffness and damping a numerical example.

Taking the rotor model in Figure 7.70 as an example, its critical speed map shows that the rotor operates between the third and the fourth criticals. Two identical 8-pole AMBs are chosen to support the rotor with dimensions in Figure 7.69. The first issue is finding the best method for determining the stiffnesses. In this case, the stiffness per bearing can be made 1000 lb/in. or 10,000 lb/in. (1.75 × 105 N/m or 1.75 × 106 ) N/m). The answer depends on rotor shock load. To take 1 g shock, this rotor of approximately 100 lbs (45 kg) moves radially 50 mi and 5 mi (1.25 mm and 0.125 mm) respectively, for the lower and higher stiffnesses. The catcher bearing is set at 10 mi (0.25 mm) away from the rotor for a designed air gap of 20 mi (.5 mm). To avoid pounding the catcher bearing when shocked, the higher stiffness is chosen. Reviewing the mode shapes at the chosen stiffness, Figure 7.71 reveals that there are sufficient relative displacements at the bearings for control of the first and second modes. The third and the fourth modes are lacking the displacement at one bearing. To help control the third mode, the displacements sensor is mounted at the outboard side of each AMB where the sensor sees more than only the center motion. The second design issue is to determine how many bending modes should be controlled. To keep the control electronics relatively simple, the frequency range with acceptable control response is limited by two factors, the inductance load and the filtering delay. It is imperative to have an adequately damped bending mode below the operating speed (the third mode in this case), because of the unbalance excitation during traversing the critical. The bending mode immediately above the operating speed (the fourth mode in this case) should be 15 to 20% away in frequency. However, it still can be excited by harmonics as the rotor is going up in speed, or by a shock load. But, less damping is required for controlling this mode. The higher bending modes normally are less likely to be excited. The rotor material damping is a source to resist the minor or occasional excitation. Oil-film bearings always provide positive damping but there is no guarantee of this for AMBs. The control current at the high critical frequency may lie behind the displacement measurement or the probe may be at the wrong side of the AMB. The AMB may become a small exciter for that mode. When it happens, a band-reject filter for the excitable mode can be inserted in series in the feedback loop to block the control at that modal frequency. For the example herein, the power amplifier and the sensor low-pass filter are assumed to have the cut-off frequencies at 500 and 5000 Hz, respectively. Applying the normalized stiffness and damping of Figure 7.69, the normalized frequency of 1.0 to 50 Hz, which is between the first and the second criticals can be read from the map. The values of ωB/K for the lower four modes range from 0.2 to 0.6, which is adequate for properly designed vibration modes. More damping can be achieved by increasing the value of Cv .

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design (a)

7-83

15 10 Coupling

Thrust disk

Inch

5

Wheel

AMB No. 1

AMB No. 2

0 –5

–10 –15 0

5

10

15

20 Inch

25

30

35

40

(b) 100,000

Inch

4th

Operating speed 15,000 rpm

10,000 3rd

2nd 1st 1,000 1,000

AMB stiffness √(K 2 + v2B2 10,000

100,000

1,000,000

Inch

FIGURE 7.70

Characteristics of the rotor model (a) Rotor model. (b) Rotor critical speed map.

7.6.3.4.2 Rigorous Dynamic Analysis After component sizing and cursory analysis, rigorous system analysis is needed to prove the expected rotor vibration behavior. Considering the fact that the AMB stiffness and damping are functions of excitation frequency, the state vector of the conventional rotor model is extended to include the state variables of the AMBs. The mathematical rotor model is the same as the conventional model including sections of shaft with specified ID, OD, length, and concentrated masses and inertias. The model for each bearing would be the bering station number, the measurement station number, and the key parameters of Figure 7.69. Using this electromechanical model, the lower four damping frequencies of the rotor running at 15,000 rpm were computed and are presented in Table 7.28. The first three modes are adequately damped because the associated log decrement values are all significantly above 0.4. The latter is a damping value generally accepted for a rotor system supported in oil-film bearings. The fourth mode has a log decrement value of 0.06 without considering the rotor material damping. It should be acceptable since it is much higher in frequency that the third mode and the operating speed. In Table 7.28 the cross-coupling stiffness produces a destabilizing seal effect at the wheel, but it only affects the first mode damping. The reason for this is that only the first mode shape has significant lateral displacement at the wheel. The rotor/AMB system can sustain a value of 4000 lb/in. (7 × 105 N/m) before becoming unstable.

© 2006 by Taylor & Francis Group, LLC

7-84

Handbook of Lubrication and Tribology AMB No. 1

1.0

AMB No. 2

0.8

Relative amplitude

0.6 0.4 0.2 –0.0 –0.2 –0.4 –0.6 –0.8 –1.0 0

FIGURE 7.71

5

10

15 20 25 Rotor length (in.)

30

35

40

Rotor critical mode shapes. Table 7.28

Damped Natural Frequencies of Forward Modes Kxy = −4000 lb/in.(−7 × 105 N/m)

Kxy = 0 Mode 1st 2nd 3rd 4th

Frequency (cpm)

Log decrement

Frequency (cpm)

Log decrement

2,318 5,263 11,154 34,156

1.05 2.20 0.89 0.06

2,374 5,623 11,144 34,156

−0.00 2.19 0.88 0.06

Note: Rotor speed = 15,000 rpm; Cross-coupling stiffness (Kxy ) at wheel.

Figure 7.72 presents an unbalance response at the third critical speed using the same electromechanical model. It indicates that the response peak is well damped and far away from the operating speed of 15,000 rpm. The peak dynamic current was calculated to be 0.45 (0-peak) at 11,500 rpm. It specifies that a current slew rate no less than 500 A/sec must be provided by the power amplifier design.

7.7 Design Considerations In practice a designer must obtain quantitative data to ascertain on the one hand whether the bearing will meet his operational requirements; and on the other hand find out what the power losses, flows, temperatures, etc. will be to properly plan the layout of the facility. In Sections 7.2 to 7.6 the graphs and tables offered values for the performance of various bearing designs. These, however, do not exhaust the information required for rational design. What is needed is some orientation as to how the various geometrical and operational parameters affect bearing operation and how to go about improving or even optimizing a given bearing design. Within their restricted space the following paragraphs should offer some guidance as to how to go about approaching this task.

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

7-85

1.40

Response (m, O-peak)

1.20

1.8 gm-cm at thrust disk 1.8 gm-cm at coupling –3.6 gm-cm at wheel

1.00

AMB No. 1

0.80 0.60 0.40 AMB No. 2 0.20 Wheel 0.00 7.00

8.00

9.00

10.00 11.00 12.00 Speed (1000 rpm)

13.00

14.00

15.00

FIGURE 7.72 Unbalance response at third critical speed. (Taken from Allaire, P.E., Li, D.F. and Choy, K.C., J. Lubr. Technol., Trans. ASME July 1980. With permission.)

7.7.1 Performance Parameters The expressions required for calculating the more important items of bearing performance are the following: Film thickness: For an aligned journal the film thickness is given by h = (h/C = 1 +  cos(θ − φ))

(7.53)

The attitude angle φ is defined as the angle between the line centers — a line passing through the centers of bearing and journal — and the load vector. When the treatment is restricted to vertical loads, φ denotes the angle between location of hmin and the vertical and therefore the importance of φ lies in that it determines the location of hmin . Sommerfeld number (load parameter): The Sommerfeld number, given by µN S= P


2 R C

(7.54a)

has traditionally been the most important parameter. However, a more convenient quantity is the inverse of S, here called the Load Parameter, given by

W =

P µn


2
2 W C C = R LDµn R

(7.54b)

where P = (W /LD) is the unit loading. What this parameter says is that any combination of P, µ, N , C, and R such as to leave the value of W unchanged, would result in the same bearing eccentricity ratio, , and attitude angle, φ.

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7-86

Handbook of Lubrication and Tribology

0.1

Hydrodynamic lubrication regime

Boundary lubrication

Low loads

f

0.001

High loads

S = (mN/P )(R/C)2

FIGURE 7.73

Behavior of friction coefficient in fluid film bearings.

Minimum film thickness: is given by

This is the smallest distance between the journal and bearing surfaces and it

h min =

hmin = (1 − ) C

(7.55)

What is normally referred to as load capacity relates to the load, W , which this hmin can support. Friction coefficient: This is the ratio between the frictional force and bearing load. It is normally expressed in the form of
 R (R/C)FT f = C W

(7.56)

The general shape of f as a function s is given in Figure 7.73. The region of sudden rise in f denotes the limit of hydrodynamic lubrication, followed by a regime of “boundary lubrication” characterized by partial contact between the mating surfaces. Power loss: This, of course, can be obtained from the value of Fτ , namely H = FT · Rω = f · W · Rω
H=

H Ho

 =

(7.57)

H [π 3 µN 2 LD 3 /c]

The quantity by which H is normalized, represents the power loss in an unloaded concentric journal bearing, that is, one in which  = 0. It is know as the Petroff Equation. Flow: An amount of lubricant, Q1 , enters the bearing at the leading edge; an amount, Qs leaks out the two sides of the bearing (one-half Qs at each side), and an amount Q2 leaves the trailing end of the pad. In most cases, since a journal bearing extends over 2π , Q2 is not discharged outside but reenters the next oil groove, so that the net amount of lubricant to be made up from an outside source is Qs . The latter is referred to as side leakage. Clearly we must always have Q1 = QS + Q2

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Principles of Gas Turbine Bearing Lubrication and Design

7-87

All of these flows are given in dimensionless form as Q=

Q (π/2)NDLC

(7.58)

the denominator representing the flow in an unloaded, concentric bearing, that is, at  = 0 (for which case QS = 0 and Q1 = Q2 ). The above flows, Q1 , QS , and Q2 are what may be called hydrodynamic flows induced by the shearing action and pressure gradients of the fluid film. QS is the minimum amount of oil to be delivered to the bearing so as to maintain a full fluid film with all its potentialities. In practice designers supply more than this required minimum, using a supply pressure ps < pa . The effect of the supply pressure, usually of the order of 10 to 30 psig, can be ignored as far as bearing hydrodynamics are concerned. Temperature rise: A bulk temperature rise can be estimated from the values of power loss and side leakage, namely T = (Tav − T1 ) = Dynamic coefficients:

H cp wQs

(7.59)

The dimensionless stiffness is given by K = (K /2µNL ) (C /R )2

while the damping coefficient reads B = (π B /µL )(C /R )3 from which the dimensional values of K and B can be obtained. The coefficients , f , H , Q 1 , Q 2 , K , and B which serve to evaluate bearing performance are obtained from solutions of the Reynolds Equation for the specific geometries and operating conditions of the various bearing designs.

7.7.2 Bearing Configuration The behavior of a bearing is naturally a function of its geometry. However, even for a given design there are a number of variables that will affect its performance. Among the more known parameters are the L /D and C /R ratios and the degree of preload. Of the less familiar ones one can cite load orientation, the geometry of the oil grooves, or the relative proportions of a bearing’s geometrical elements. 7.7.2.1 Journal Bearings Although one often hears about the use of full, that is 360◦ arc bearings, it is very rarely that such sleeves are employed in machinery. Most journal bearings consist of two or more pads separated by horizontal oil grooves making them in fact partial bearings, used either singly or in tandem. The number and distribution of these angular pads on bearing performance is one of the more important considerations in bearing design. 7.7.2.1.1 Partial Bearings Whenever a single pad of an angular extent β < 2π is used, it is called a partial bearing. When β is very small its load capacity is low, as illustrated in Figures 7.74 and Figure 7.75. However, soon a limit is reached at about β = 140◦ beyond which no further gains are registered. The reason for this asymptotic behavior is due to oil cavitation at the trailing end of the pad where the pressures decrease close to or even below ambient pressure. This, if a partial bearing is used there is no need to go beyond a 140◦ arc. The effect of temperature in partial bearings is a combination of two phenomena. The higher the arc the longer the

© 2006 by Taylor & Francis Group, LLC

7-88

Handbook of Lubrication and Tribology W

O Or

100

1 Dimensionless load capacity, W

10 1/2 1/4

1

1.0 1/2

1/4 0.10

e = 0.2 e = 0.8 Nos. refer to (L/D) 0.01

FIGURE 7.74

20

60

100 140 180 Bearing arc b°

200

360

Effect of bearing arc on load capacity.

dissipation path and the higher the temperatures; however, a longer arc produces thicker films and thus less heating. Consequently, as shown in Figure 7.76, a crossover point occurs; at high loads low values of β are preferred, if low T s are desired; at low loads a longer arc is preferred. 7.7.2.1.2 Grooved Bearings Partial bearings are not used extensively. The most common designs are grooved bearings which consist of a number of pads arranged in tandem by cutting axial oil grooves in the 360◦ circumference. There is a great variety of such designs, the most common being a 2-pad bearing with two grooves at the horizontal split. Others may have 3, 4, or 6 grooves forming the same number of individual pads. The more the number of grooves the lower the load capacity, as shown in Figure 7.77 and 7.78. Thus, if load capacity is the primary objective, a 2-groove bearing is best; however, those with a larger number of grooves are somewhat more stable. Related to the above is the fact that any hole or disruption in the bearing surface will reduce the load capacity. Figure 7.79 shows the effects on the pressure profile of cutting a slit or circular hole in the loaded part of a bearing. The larger the incursion, the more drastic the reduction in the hydrodynamic pressures which translates directly into reduced load capacity. 7.7.2.1.3 Tilting Pad Bearings The primary characteristic of this family of bearings is that the individual pads are not fixed but are pivot supported so that during operation not only does the journal move but so do the pads and each in a different fashion. A general picture of a tilting 3-pad bearing is shown in Figure 7.80. The structural and

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Principles of Gas Turbine Bearing Lubrication and Design

7-89

1.0 0.9 Minimum film thickness (hmin/C)

b = 360 0.8 b = 180

0.7 0.6 0.5 0.4 0.3

b = 120 0.2 b = 60

0.1 0 0.001

0.01

0.1 Load capacity, (mNP) (R/C)2

1.0

10.

FIGURE 7.75 Effect of bearing arc on value of hmin . (Taken from Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. With permission.)

103 8 6

L/D = 1

4

Temperature rise (wc D T/P)

2 102 8 6 4

b = 60° b = 120°

2 10 8 6 4 2

b = 180° b = 360°

1

0.01 0.02

0.06 0.1

0.2

(m N/P)

0.6 1.0

2.0

6.0 10.

(R/C)2

FIGURE 7.76 Effect of bearing arc on temperature rise. (Taken from Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. With permission.)

© 2006 by Taylor & Francis Group, LLC

7-90

Handbook of Lubrication and Tribology fL = 0 +fL –fL L/D = 1 e = 0.6

10 9

1

2

3

Load capacity, W

1 8 7 2

6 3

5 4 3 2 1 0 –60 –50 –40 –30 –20 –10 0

10 20 30 40 50 60 70 80 90 100 110 120 Load angle fL°

FIGURE 7.77 Load capacity of grooved bearings. (Taken from Pinkus, O., J. Lubr. Technol., Trans. ASME, Oct. 1975. With permission.) 1.0 0.8 0.7 0.6 0.5

drica

cylin

ial cy

e

4 Ax

Plain

0.3

5 = 0. L/D 5 brg, = 0. ove , L/D l gro brg Axia ical 1 r d L/D = ylin brg, in c Pla rical n li =1

0.4

l brg

0.2

, L/D

0.1 0.01

0.1

1.0 S

FIGURE 7.78

Comparisons of 2- and 4-axial bearings.

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10

Principles of Gas Turbine Bearing Lubrication and Design

7-91

1.0 L/D = 1/2; e = 0.5

L/D = 1/2; e = 0.5

No slot

No hole 0.05 L

5 L

L0

d 0.05

0.20 Extent of hole at C.L. Mean axial pressure, P

0.60

7.2°

0.40

Values (L/L0)

0.02

0.04

1.0

0.5 0.60

v

0.80 Values (d/L)

Load 54°

90°

Pmax 0.1 126°

162°

54°

199°

90°

126° Position from hmax

Position from hmax

FIGURE 7.79 Effect of a slot and a hole on hydrodynamic pressure.

Pivot Pivot clearance circle

Pad R-Cm e

v

R

Shaft

bp

FIGURE 7.80 A tilting 3-pad journal bearing.

© 2006 by Taylor & Francis Group, LLC

b

162°

199°

7-92

Handbook of Lubrication and Tribology Table 7.29 Relative Load Capacity and 3-Pad and 5-Pad Bearings Values of m On-pivot load

Load between pivots

W

3 Pads

5 Pads

3 Pads

5 Pads

20 40 60 80 100

0.72 0.79 0.80 0.82 0.83

0.82 0.87 0.90 0.92 0.93

1.42 1.58 1.60 1.61 1.62

0.97 1.05 1.09 1.10 1.11

Source: Pinkus, O., J. Tribol., Oct. 1986.

analytical complexities of these bearings are more than compensated by their great reliability and the fact that they have no rival in their stability characteristics. The number of possible design parameters and operating modes in a tilting pad bearing is very large. Some of them are discussed below. 1. Number of pads: Table 7.29 gives a comparison of a 3-pad vs. a 5-pad centrally pivoted bearing having zero preload. When the load is in line with the pivot the 3-pad design has a higher load capacity but the reverse is true when the load direction is between the pads. For loads of engineering interest the 5-pad design consumes less power. 2. Pivot location: In order to assure two-directional rotation and for ease of assembly, most tilting pad bearings are centrally pivoted. However, a 10 or 15% displacement of the pivot in either direction would not significantly alter the general performance, a slight preference being a downward shift. 3. Preload: From many standpoints a high preload is desirable. Its effect on preventing the scraping of the top pads has been discussed previously and from this standpoint a m of at least 0.5 is required. High preloads also yield higher stiffness and damping. However, the penalty is that the film thickness over the pivot and often also the absolute hmin is reduced. Likewise, the power losses and temperatures rise with an increase in preload. 4. Mode of loading: In general the shaft eccentricity will be lower when loaded over the pivot. It is characteristic of tilting padis bearings that regardless of whether the load vector is over the pivot or between the pads the locus of shaft center is along a vertical line, which has a direct beneficial effect on stability. Results for the two modes of loading on stiffness and damping are given in Figure 7.81 for a bearing of zero preload. As seen, both the spring and damping coefficients are lower for the between-pads mode of loading. 7.7.2.1.4 Oil-Ring Bearings As pointed out previously, oil-ring bearings operate under starved conditions. It is thus the main task of the designer to find ways to increase as much as possible the amount of oil delivered to the bearing surface. Some of the important parameters that play a role in accomplishing it are geometry shape of contact surface, weight, the material, and the size of the ring relative to the shaft. In an experimental study a series of rings portrayed in Table 7.30 was tested with the purpose of both increasing the flow of lubricant and of extending the regime of stable ring operation. The conclusions reached were as follows: 1. An optimum ring shape is one with a quasi-trapezoidal cross section and a series of straight teeth at the contact surface shown in Table 7.30 as Ring No. 2. 2. The best ring material is bronze with a weight of 23 N per meter of ring circumference. 3. For bearing diameters in excess of 6 in. dual rings are recommended. 4. An anchored spring leaf inserted between the ring and journal raises the amount of oil delivery and extends the ring’s region of stable operation. One such stabilizer is shown in Figure 7.82.

© 2006 by Taylor & Francis Group, LLC

103

10

1

102

1

Cv Bxx /W

10

CKxx /W

10–1

Cv Byy /W

CKyy /W

1

10–2

10–3

10–1 10–1

1 10 S = (m NDL/W) (R/C)2

102

CWMCR/(mLD(R/C)2)2

Critical mass

102 CK/W; Cv B/W

10

CK/W; Cv B/W

103

7-93

Critical mass Cv Bxx /W

10

CKxx /W

1

10–1

10–2

CWMCR/(mLD(R/C)2)2

Principles of Gas Turbine Bearing Lubrication and Design

10–3

10–1 10–1

1 10 S = (m NDL/W) (R/C)2

102

FIGURE 7.81 Effect of mode of loading on bearing stability in a 4-pad tilting pad bearing. (a) Load between pivots. (b) Load over pivots.

7.7.2.1.5 Load Angle Bearing loads are usually directed midway of a pad or between grooves. However, improved performance can be obtained by shifting the load vector toward the trailing edge of the bearing pad. A comprehensive mapping of the effects of shifting the load vector around the circumference of a 2-groove bearing is shown in Figure 7.83. Normally the load would be straight down, that is along φL = 0. However, as seen in the figure by moving the load toward the trailing edge, improved performance is obtained for the entire range of bearing operation. At low loads an optimum occurs at a load of φL = 10; at high loads the value is some 30. The lowest load capacity would occur at a load angle of 60◦ from the midway point. Supplementary data is given in Table 7.31 where it is seen that the worst angular position results in a load capacity reduction of 70.70%. Similar data for a 3-groove bearing has been given in Figure 7.77. Achievement of an optimum bearing position requires no special effort. It is sufficient to rotate the bearing in the housing the required 10 to 30◦ to obtain this. Attention should only be given to the oil delivery path since now the oil grooves would no longer be at the horizontal split. This can be taken care of by cutting a short oil-supply channel on the outside of the bearing shell. 7.7.2.1.6 Misalignment It was pointed out in an earlier section that an overhung impeller will cause bearing misalignment. A full treatment of misalignment would exceed the available space here but the qualitative consequences should be pointed out. As shown in Figure 7.84 in severe misalignment the journal at one end may find itself in the upper half of the bearing even though the load is downward. As a consequence, a fluid film and hydrodynamic pressures may develop in both the lower and upper portions of the bearing. Stretching from the end where the hydrodynamic film is at the bottom, this film will wrap itself in helical fashion around the entire bearing circumference. In all cases the load capacity, that is the value of h for the imposed load, will be drastically reduced. 7.7.2.2 Thrust Bearings Unit loads in thrust bearings are higher than in journal bearings and consequently their hmin will be smaller. But it should also be realized that, except for a bearing with a flat at the end, hmin in thrust bearings occurs not along a line as in journal bearings but at a point, namely the outer downstream edge

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7-94

Handbook of Lubrication and Tribology Table 7.30 Oil ring configuration

Ring No.

T Split T-Section Brass Ring Made from Rolled Stock Fastened Together

1

3.4

3.4 C

11.7

A

7.35

E

19.9 "

Split Trapezodial Section Machined from Bronze SAE660 "=30E,$b=0E

2

Unit Weighta (N/m)

Cross section (mm)

Description

3.2

1.3

14.3

D

23.6

C

2.4

E

B 25.4

Split Trapezodial Section Machined from Bronze SAE660"=30E,$b=90E

3

1.7 24.9 B $

Split Trapezodial Section Machined from Bronze SAE660 "=30E,$b=45Eor135E

4

3.0

1.7 24.3 B 3.0

$ Split and Relieved, Trapezodial Section Machined from Bronze SAE 660 "=30E,$b=0E

5

aUnit

1.5

18.0 B L

22.9

5.2

weight equalsring weight/circumferential length. $ is measured from the direction of ring rotation.

bAngle

Source: Heshmat, H. and Pinkus, O. “Experimental Study of Stable High-Speed Oil Rings,” J. Trib. Trans. ASME, 107, 1985, 14–22.

of the pad. This point is also where Tmax will occur and again it will be higher than in journal bearings. This is due to the smallness of hmin but also to the higher linear velocities of the runner at the outer radius of the pad. 7.7.2.2.1 Tapered Land Bearings A conventional tapered land bearing was shown in Figure 7.31. There are three parameters here; the taper (h1 − h2 ), the pad arc β, and the (L/R2 ) ratio. The angular extent also determines the number of pads in

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Principles of Gas Turbine Bearing Lubrication and Design (a)

7-95

Ring

NR

Oil film Journal N

(b)

Ring backward swing Ring stable position

A

Ring

Ring forward swing

N

View A–A

FIGURE 7.82 Configuration of oil-ring stabilizer. (a) Spring stabilizer. (b) Stabilizer and ring positions during oscillations. Table 7.31 Effect of Load Angle on Load Capacity in Conventional 2-Groove Bearing Worst Condition 

W at Worst φL W at φL = 0

L /D



W at φL = 0

φL

W

0.5

0.6 0.95

3.1 83

50 55

1.05 14.5

0.32 0.175

1.0

0.6 0.95

8 115

52 60

2 20

0.25 0.17



a thrust bearing. Table 7.32 shows the results of an optimization study giving the values of (h1 − h2 ) and β for the entire range of (L /R2 ) ratios. From this an optimum set of design parameters can be obtained for a particular application. It is worth noting that in general the optimum configuration is that which yields nearly square bearing pads. An improved version of a plain tapered land bearing is one with a flat surface at the trailing end, as shown in Figure 7.32. The additional merit of this design is that upon starting and stopping the runner rides on a flat surface reducing wear. Here a new parameter is the ratio of the tapered to the flat portion. The plot in Figure 7.85 shows such a variation from 60 to 100% taper, the latter being the tapered land bearing discussed previously. The load capacity peaks at a taper value of about 80% of the pad arc, that is the tapered portion should be four times that of the flat. Interestingly the value of (power loss/load capacity) achieves a minimum at the same point. 7.7.2.2.2 Misalignment In properly operating thrust bearings the load carried by each pad is the same. When the shaft and consequently the runner is misaligned, this is no longer true and some of the pads are much more heavily

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7-96

Handbook of Lubrication and Tribology

1000

Location of uG1 500 groove

Central loading

Location of uG2 0.986 groove 0.98

200

0.95

100 0.9

Load capacity, W

50

0.8 20 10

0.6

L/D = 1

5

+fL –fL

e = 0.4 2

Load 1

v uG1

0.5

uG2

0.2 0.1 –80 –60 –40 –20

FIGURE 7.83

0

20 40 60 80 Load angle fL, degrees

100

120

140

160

Effect of load angle on load capacity in 2-groove bearing.

Table 7.32 Optimum Pad Arrangement |25| L /R2 1/3

1/2

2/3

h1 /δ0

β, deg

Number of pads

1 1/2 1/4 1/8 1 1/2 1/4 1/8 1 1/2 1/4 1/3

10 >10 9 8 8 7 6 5 6 5 4 4

Source: Pinkus, O., Trans. ASME, Ser.D., 81, 1959.

loaded than the other. A pictorial representation of this situation is given in Figure 7.86. As seen, the loads carried by the heavily loaded pads as well as their maximum temperatures can be ten times as high as the ones located opposite them where the runner is furthest from the pads. The values of h in the two sets of pads will be of the same ratio. The span of severity of bearing operation goes up with the number of pads used in the misaligned bearing. Thus, if misalignment is expected one should not use more than four to six pads.

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Principles of Gas Turbine Bearing Lubrication and Design

7-97

F2

O

O⬘2

O⬘1

F1 N F2 Viewed from 2

Hydrodynamic film

Hydrodynamic film

F1

FIGURE 7.84

Hydrodynamic forces and films under misalignment.

7.7.2.2.3 Hydrostatic Bearings In a conventional hydrostatic bearing portrayed in Figure 7.4 the load capacity is given by W =

π R22 (po − pa )[1 − (R1 /R2 )2 ] 2n (Re /R1 )

There are therefore two parameters that determine the level of W ; (po − pa ) and (R2 /R1 ). The variation of load with these quantities is shown in Figure 7.87. As seen no optimum for load capacity occurs; it rises with p and drops with a rise in (R1 /R2 ). A minimum occurs in the power loss but power loss in a hydrostatic bearing is not of great concern and, when it is, it is due not to bearing geometry but to the onset of turbulence in the fluid.

7.7.3 Qualitative Guidelines In selecting design parameters it must be kept in mind that the choice often depends on the size of the bearing. Small bearings, less than 2 in. in diameter, can tolerate relatively lower values of hmin , higher unit loads of P, and operate close to isothermal conditions, whereas larger bearings require larger values of hmin , lower values of P, and tend to run close to adiabatic conditions. On the other hand (C/R) ratios must be higher for small bearings. With this as an introduction Table 7.33 gives some typical design practices in the field of journal bearings. The performance characteristics of one’s design should fall somewhere within the range of values listed in the table.

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7-98

Handbook of Lubrication and Tribology 150 D(DT )max 140

130 W = 0.0286 120 DW 110 DH 100

90 D Qs

D H/D W 80

70 Tapered land thrust bearing 60

b = 32°; (L/R2) = 2 h11 = 3; d21 = 0;

50 Rei = 700; E = 3.3 60

FIGURE 7.85

70

80 b%

90

100

Effect of extent of taper or flat. Table 7.33

Typical Design Limits for Journal Bearings

Minimum film thickness Temperature rise Maximum temperature Loads L/D Ratio C/R Ratio Preload, m Bearing arcs Inlet oil temperature, T

0.001–0.01 in. (0.0025–0.25 mm) Up to 80◦ F (27◦ C) (on babbitt) Up to 300◦ F (150◦ C) (on babbitt) 500 psi (3.4 Mpa) 0.25–1.0 0.001–0.002 0.25–0.75 150–60◦ for fixed pad 80–30◦ for tilting pad 80–130◦ F (27–55◦ C)

Nearly all the bearing data given in the present write-up are for bearings operating under laminar conditions. Should turbulence set in, the operating characteristics will change. One may expect turbulence when the bearing Reynolds Number reaches a value between 750 and 1500. The higher the Reynolds Number the more intense will be the effect of turbulence. Table 7.34 shows what will be the impact of the turbulent regime on the major items of bearing operation. In a more comprehensive way Table 7.35 provides a guide in which direction design modifications should head in order to ameliorate unsatisfactory results in a chosen design. Finally, Table 7.36 offers a cursory look at the relative advantages and disadvantages in choosing journal bearings of different designs.

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Principles of Gas Turbine Bearing Lubrication and Design

7-99

Pressure protites in 12-pad bearing: ax = 0 .75, ay = Q, E = 12.7 9

Multipad thrust bearings E =1.27 ax = 0.75 ay = 0 n = 12 Pmax/P n=6

8

Multipad thrust bearings E =12.7 ay = 0 ax = 0.75

3.6

7 3.2 6

n = 12

2.8

DTmax

n=8

2.4

5

2.0 WT

4

n=6

1.8 3 1.2 2 Q2

0.8

hmin

0.4

1 M

0

0 0

p u Performance of individual pads in multipad thrust bearing

FIGURE 7.86 permission.)

2p

0

Bearing pad number

n

Pad loadings in multipad thrust bearings

Effects of misalignment in thrust bearings. (Taken from Pinkus, O. ASME, Series D, 83, 1961. With

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7-100

Handbook of Lubrication and Tribology

Q

Q H

W

W

H

0

1 (R1/R2) h

R1

R2

FIGURE 7.87

Performance of incompressible hydrostatic bearing.

Table 7.34 Effect of Turbulence on Bearing Performance Reynolds Number a Parameter Re = ρRw h /µ Regimes • Re < 750 Laminar • 750 < Re < 1,500 Transition • Re > 15,000 Turbulence Item

Effect

Load capacity Oil flow Power loss Temperatures Stiffness and damping

⊕  ⊕ ⊕ ⊕ or 

⊕ — Increase;  — Decease

7.8 Advanced Bearing and Seal Applications Driven by the goals of increased power density and higher performance, advanced turbomachinery is operating at ever higher speeds and temperatures. Some of these advanced designs will require equally advanced bearing systems that do not suffer from the temperature limitations of liquid-lubricated bearings or the limited life of REBs when operating at extreme conditions. One such oil-free bearing, which has been applied successfully in advanced, high-speed machinery, is the air-lubricated, CFB as described in Section 7.6.2 (see, e.g., [38–41]). A recent summary of progress in the state of the art for foil bearings is presented in Reference 42. These bearings have been applied to a wide range of high-speed machinery, with operating environments ranging from cryogenic to temperatures in excess of 1500◦ F and speeds in excess of 700,000 RPM.

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Principles of Gas Turbine Bearing Lubrication and Design Table 7.35

7-101

Effect of Design Parameters on Journal Bearings Performance

Code: (+) means increase magnitude of parameter to achieve effect in left-hand column. (-) means decrease magnitude of parameter to achieve effect in left-hand column. Example: To decrease temperature rise, one or more of the following can be done: decrease L/D ratio, increase C/R, decrease oil viscosity, etc. Objective

L/D

C/R

To decrease temperature rise (−) (+) To reduce power loss (−) (+) or (−) (+) (+) To reduce Tmax To increase oil flow (+) (+) To improve stability (−) (−) To increase load capacity (+) or (−) (−) To avoid turbulence (+) or (−) (−) Stability (+) (−)

Geometry

Viscosity

Preload, m

Arc, β

(−) (−) (−) (−) (+) (−) (+) (+)

(−) (−) (+) or (−) (+) (−) (−) (−) (+)

Elliptical Circular Elliptical Elliptical

(−) (−) (−) (−) a (+) or (−) Circular (+) 3-lobe (+) Tilting pad (+) or (−)

Supply Oil Pressure, ps (+)b (−) No effect (+) No effect No effect (−) No effect

a The stability of a journal bearing increases in the following order: circular, pressure, elliptical, 3-lobe, tilting pad. b Apparent effect only.

Table 7.36

Characteristics of Various Journal Bearings Journal Bearing Summary Table

Bearing type

Advantages

Disadvantages

Axial groove

1. Easy to make 2. Low cost

1. Subject to oil whirl

Elliptical

1. Easy to make

3- and 4-lobe (tapered land, etc)

2. Low cost 3. Good damping at critical speeds 1. Good suppression of whirl

1. Subject to oil whirl at high speeds 2. Load direction must be known

2. Subject to whirl at high speeds

Pressure dam (Single dam)

2. Overall good performance 3. Moderate cost 1. Good suppression of whirl 2. Low cost

2. Dam may be subject to wear or build up over time 3. Load direction must be known

Hydrostatic

Tilting pad

3. Good damping at critical speeds 4. Easy to make 1. Good suppression of oil whirl 2. Wide range of design parameters. 3. Moderate cost 1. Will not cause whirl (no cross coupling) 2. Wide range of design parameters

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1. Some types can be expensive to make properly

1. Goes unstable with little warning

1. Poor damping at critical speeds

Comments Round bearings are nearly always “crushed” to make elliptical or multi-lobe Probably most widely used bearing at low or moderate speeds

Currently used by some manufacturers as standard bearing design

Very popular with petro-chemical industry. Easy to convert elliptical over to pressure dam

Generally high stiffness properties used for high precision rotors

2. Requires careful design 3. Requires high pressure lubricant supply 1. High cost

2. Requires careful design 3. Poor damping at critical speeds 4. Hard to determine actual clearances 5. High horsepower loss

Widely used bearing to stabilize machines with subsynchronous nonbearing excitations

7-102

Handbook of Lubrication and Tribology

Khydro

Hydrodynamic component of stiffness

Kstruct Compliant component of stiffness

Bhydro

Bstruct

FIGURE 7.88

Hydrodynamic component of damping

Compliant component of damping

Foil bearing operating mechanisms.

As shown in Figure 7.88, the essential feature of such a bearing is its twofold mechanism of imparting stiffness and damping to the system. One is via the geometry and materials of the complaint support structure. The other is due to the hydrodynamic film between journal and top foil. The compliant support can consist of one or more corrugated bump foils, which offer great flexibility in obtaining dynamic characteristics geared to a specific system. Additional flexibility is afforded through construction as a single top foil, or a multipad bearing with multiple top foil segments. The complaint construction allows the foils to be forced radially outward as speed increases, forming a converging wedge. This converging wedge becomes more pronounced with increasing speed and load, thereby increasing load capacity. The compliant foil surface readily accommodates itself to rotor centrifugal and thermal growth, as well as thermal and mechanical deformations of the bearing housing. These bearings have also demonstrated good performance under shock loading [43]. From a historical perspective there has been a considerable advance in CFB technology over the last decade. The capabilities of advanced designs now meet the requirements for advanced applications, with unit loadings in excess of 100 psi (689 kPa) [40] for journal bearings, 85 psi (586 kPa) for thrust foil bearings, and adequate damping to allow successful operation above the first system bending critical speed [44–46]. However, the issue of the scalability of early designs, which were typically 25 to 50 mm, to the much larger sizes required by turbomachinery such as gas and steam turbines had not been widely addressed until the late 1990s. The largest bearing discussed in the open literature, which was located by the authors, is the 89 mm diameter bearing described in Reference 38. This bearing was presented as having an ability to support a steady-state operating load of 23.5 lb (109 N), and a dynamic load of 335 lb (1490 N). These loads were well below that required by even most small gas turbine engines. However, recent efforts by Heshmat et al. 40–46 have been completed that have substantially extended the applicability of foil bearings to bearing sizes and loads that are substantially higher than previously demonstrated. In particular, static loads of 950 lb (4200N) have been supported by a 100 mm diameter foil journal bearing operating at a speed of 22,000 rpm. Figure 7.89 shows the measured and predicted load-carrying capability of two modern fourth generation foil journal bearings as a function of speed

© 2006 by Taylor & Francis Group, LLC

Principles of Gas Turbine Bearing Lubrication and Design

2600

7-103

150 mm bearing

2400 2200 2000

Applied load (lb)

1800 1600 1400 1200 1000 800 100 mm bearing 600 400 200 0 0

FIGURE 7.89

5

10 15 20 Speed (krpm)

25

30

35

Load-carrying capability of 100 and 150 mm diameter foil journal bearings.

capable of supporting static loads in excess of 2500 lbs. Figure 7.90 shows that a new class of oil-free foil thrust bearings has also been developed that are capable of operating at surface velocities above 1200 ft/sec (speeds to 80,000 rpm for the specific design tested) and has demonstrated a load capacity of over 80 psi1 . This demonstrated performance is more than three times the capacity of available thrust foil bearings just a few years ago. The features that have made it possible to significantly enhance the load-carrying capability of our compliant foil thrust bearings are: 1. 2. 3. 4. 5.

A hydrodynamic pressure profile which is more uniform over the thrust pad surface. Spatially variable stiffness to accommodate axial motion and misalignment of the rotating part. Enhanced thermal management capabilities. Unique surface coatings, which can accommodate thermal demands. Light weight and low parts count for high frequency response and reliability.

These tremendous recent advances point to the very real viability of applying foil thrust bearings in gas turbine engines. Figure 7.91 and Figure 7.92 are examples of very large compliant foil journal and thrust bearings that are being fabricated today. Besides high-load capacity, additional developments have been made in the high-temperature operating of foil bearings due to improved high-temperature coatings. As seen in Figure 7.93, a 150 mm diameter foil journal bearing is being operated at temperatures above 1200◦ F and speeds above 20,000 rpm. At these operating conditions the shaft journal expansion exceeds 0.030 in. The inherent design features of the compliant surface of the foil bearing are such that it readily accommodates this substantial rotor growth. Figure 7.94 shows potential application locations 1 Where

load capacity is defined as the maximum load carried by the bearing divided by its projected area.

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology G G G

80

Achievement 2004 70 Achievement 2002 60

Load capacity (psi)

50 Bearing OD 3.82 in. Coating: Korolon 800 40

30 Achievement 2001 20 Baseline 1994 10

0 0 Speed (krpm)/(ft/sec)

Uncoated baseline MiTi Proprietary and confidential 10 170

20 340

30 510

40 680

FIGURE 7.90

Foil thrust bearing developments.

FIGURE 7.91

Large diameter, high-load capacity CFBs.

© 2006 by Taylor & Francis Group, LLC

50 850

60 1020

70 1190

80 1360

90 1530

Principles of Gas Turbine Bearing Lubrication and Design

FIGURE 7.92

7-105

Large diameter, high-load capacity thrust foil bearing.

for oil-free bearings and seals in advanced gas turbines. In certain applications where rotor loads are very high, compliant foil or hybrid foil/magnetic bearings may be considered. Many other oil-free hybrid combinations are also under consideration. This foil bearing technology base was instrumental in the successful test of a WJ24-8 240 lb thrust turbojet engine with a foil bearing directly behind the turbine. The foil bearing, tail cone, and rotor used in this test are all shown in Figure 7.95. The engine ran flawlessly, vibrations were low, and engine fuel consumption was reduced. Posttest inspection of the engine and bearings revealed them to be in perfect condition. Besides the foil bearing employed in the turbojet testing, foil bearings ranging from 6 mm in. diameter to 150 mm in diameter (see Figure 7.96 through Figure 7.98) have been developed and tested. What is interesting to note from Figure 7.96, is that as bearing size has decreased to 15 mm in. diameter and smaller, the predicted power loss due to viscous shearing of the air has been reduced. This reduction in power loss is due primarily to the fact that the static loading on the bearing is approximately one order of magnitude smaller than the larger bearings. This low power loss, which is already substantially lower than ball bearings, will be instrumental in the overall engine system thermal management for mesoscopic and MEMS class systems. One set of the small foil bearings indicated in Figure 7.96, are shown in Figure 7.97 (a) and (b). The 15 mm journal bearings and 35 mm diameter thrust bearings shown in (a) have been successfully tested at speeds in excess of 258,000 rpm and been subjected to shock loads in excess of 90 g without failure.

7.8.1 Rotor-Bearing Dynamics and Engine Integration Due to its principal of operation, the behavior of a CFB is strongly coupled to the dynamics of the interfacing rotor. Modifications to the rotor often impact the behavior of the bearing and vice versa. The design of a CFB and oil-free rotor system is therefore an iterative process that cycles between component

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Largest foil bearing ever built in the world

150 mm foil bearing under 400 Ib radfial load 2,500

20,000

Speed (20,000 rpm)

Operated to 27,500 rpm (4.05 MDN) Avg. housing temp ~ 1500°F+

1,500

15,000

Foil brg temp ~ 1200°F 10,000

1,000

Speed rpm

Temperature °F

2,000

Maximum ambient temperature to date 1500°F 120 lb rotor weight

5,000 500

00 4:

00 4:

:0 11

00 :4

4: :2

10

00 10

:0

4:

00 10

:4

4:

00 09

4:

00 :2

4: 09

:0 09

08

:4

4:

00

0

Run time

FIGURE 7.93

150 mm diameter foil journal bearing tested at elevated temperature.

Advabced bearing and seal technologies for next generation gas turbine engines Hybird foilmagnetic bearing

High performance foil bearing High temperature compliant foil seal

FIGURE 7.94 Advanced oil-free bearings and seals for gas turbine engines.

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Principles of Gas Turbine Bearing Lubrication and Design

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FIGURE 7.95 WJ24-8 rotor, tailcone and MiTi® foil bearing after engine test.

Power loss (Hp)

1.0

15 mm diameter bearings (predicted)

0.5 2.339E-009* × (1.76) 6 mm diameter bearings (predicted)

0.0 0

FIGURE 7.96

10,000

20,000 30,000 40,000 Surface velocity (ft/min)

50,000

60,000

70,000

Power loss vs. speed for foil bearings from 6 to 200 mm in diameter.

and system level performance assessment. A generalization of this process is illustrated in Figure 7.99. The initial steps involve the creation of a rotordynamic model using assumed rotor and bearing geometries. The model is then used to solve for static bearing loads, and to generate a critical speed map which is, essentially, a parametric study of the system critical speeds as a function of bearing stiffness. The functional requirements of the foil bearings are generated from this data. The effort then shifts to the component level wherein MiTi®’s specialized computer codes are used to predict the speed-dependent stiffness, damping, and power loss. The data yielded by this analysis constitutes a preliminary bearing design. Focus is then shifted back to system behavior by applying the CFB stiffness and damping coefficients to the existing rotordynamic model and investigating the response. The goal at this stage is, generally speaking, to obtain acceptable bearing loads and stable operation throughout the run range. If such results are achieved then the bearing design is complete, excepting optimization and enhancement. More often than not, however, additional design iterations are required, either on the rotor or bearing or both.

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Handbook of Lubrication and Tribology

(a)

(b)

FIGURE 7.97

Miniature foil bearings tested to over (a) 250,000 rpm and (v) 700,000 rpm.

FIGURE 7.98

MiTi® foil bearings from 15 mm to 150 mm diameter.

The prediction of CFB performance involves a blend of computational techniques and empirical data drawn from past experience. In the case of the former, a sophisticated coupled structural-hydrodynamic computer code is needed to analyze the hydrodynamics within the bearing at various speed, load, and temperature conditions. The information gained from the analyses includes, but is not limited to, the bearing’s power loss, eccentricity, and speed-dependent stiffness coefficients. MiTi®’s experience with CFBs has shown that the predicted values agree very well with experimental data. Damping, however, is more difficult to predict. Although analytical tools do exist to forecast damping, experience in this case has shown that the best predictions are made using past experimental data and engineering judgment. Although damping is often simplified to pure viscous damping in dynamic models, the available data clearly indicates that such assumptions are not valid in CFB rotor systems. The difficulty in obtaining analytical values stems from the fact that damping in CFBs is essentially a friction phenomenon and that the physics underlying friction are not well understood. 7.8.1.1 Foil Bearing Analysis Like the rotor-bearing system analysis the foil bearing analysis is also iterative in nature combining both structural and hydrodynamic compressible gas solutions. In certain critical applications, a finite element

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Principles of Gas Turbine Bearing Lubrication and Design

7-109

Assume Assumerotor rotorconfiguration configuration Available brg space Assume Assumebearing bearinggeometry geometry Brg geometry Create Createrotordynamic rotordynamicmodel model Finite element model

Static analysis

Critical Critical speed speed map map

Required stiffness

Brg loads

Hydrodynamic bearing analysis Dynamic brg coefficients (k, b) Rotordynamic Rotordynamicanalysis analysiswith with dynamicbearing bearingcoefficients coefficients dynamic Predicted rotor behavior

Viable design?

NO

YES DONE

FIGURE 7.99

Rotor-bearing system design process.

analysis (FEA) of the top foil to assess the expected stress and deformations may be required. With proper design, the stresses in the foils may be held well below the fatigue endurance limit levels of the foil material. For example MiTi strives to maintain stresses less than 2200 psi. Such an FEA is performed only if extremely high loading conditions are expected. Having already established the influence of the top foil on bearing performance through FEA, MiTi® does not routinely require the FEA to be completed unless extreme loads, speeds, and temperatures are expected, and where the contribution of the top foil may be crucial to the successful operation of the rotor-bearing system. However, most bearing systems may be successfully completed by conducting parametric design tradeoff studies of key foil bearing design parameters, such as those shown in Figure 7.100, including diameter, length, operating clearance, and structural stiffness, which is based upon the number of layers of bumps, thickness of the bump material, change in material properties with temperature, and so on. The significant number of parameters available to the designer in meeting the requirements for rotor centrifugal and thermal growth while yielding the desired stiffness characteristics offer considerable flexibility but also challenges due to the large number of parameters one has at their disposal. As such a well-defined design procedure is needed that includes the entire rotor- bearing system as noted in Figure 7.99. For example, once static rotor load, speed, and ambient temperature are known, the MiTi® procedure is to establish a nominal bearing geometry based on the length and diameter of the bearing that will maintain a low load on the foil and coating so that the

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Handbook of Lubrication and Tribology

Speed

Pressure

Static load

Dynamic load

Temperature

Thermal gradient

Bearing length

Bearing diameter

Clearance

Preload

Bump stiffness (thickness, No. of layers, No. of bumps)

Number of bearing pads (1, 3, 5, 7)

Material properties

Gas properties

Hydrodynamic pressure profile

(a)

Structural stiffness elements Gas film

Coulomb friction

Pressure (psi)

30 20

Lower

Upper

Interface coating

Bearing support/mount struture

10 (b) 0

Pressure

8

Ax

Low load condition Rotor load Original position Deflected position

100

6

ia

ld

ire

4

ct

FIGURE 7.100

io

n

2

25

Coulomb 75 frication tion c e 50 ir al d renti fe m Circu MiTi Proprietary information

High load condition

Foil bearing design parameters resulting in coupled hydrodynamic and structural stiffness.

rotor will lift off and become airborne as quickly as possible. Low-speed lift off is essential for long life of the high-temperature coating and to minimize starting sizing requirements. Once the bearing geometry has been established an initial structural stiffness is defined and the coupled structural and hydrodynamic analysis is run to determine hydrodynamic pressures and hence load-carrying capability, power loss, and stiffness coefficients. Due to the compliance of the bearing surface, an iterative solution is needed to converge on a pressure profile that matches with the deformed foil surface. 7.8.1.2 Hybrid Bearings As noted above, these recent developments are making CFB technology suitable for use as main rotor support for gas turbine engine rotors weighing as little as a few grams to more than a thousand pounds. CFBs with load capacities and damping significantly greater than previously demonstrated have been developed and combined with magnetic bearings to support rotor loads in excess of 1350 lb for a 100 mm diameter bearing. The hybrid foil/magnetic bearing combines two oil-free bearing technologies to take advantage of the strengths of each. Foil bearings have good load-carrying ability and transient shock response at high shaft speeds. Magnetic bearings provide nearly constant load carrying ability over the operating speed range, but are susceptible to overload during transient events, and have potentially catastrophic failure modes if power, sensor, or winding failures occur. Unlike more conventional magnetic bearing applications, the hybrid foil/magnetic bearing does not require a separate backup bearing. In the hybrid bearing, the foil bearing component provides the transient/failure protection, as well as significantly increases the load capacity for a given bearing system size and weight. For example, a hybrid foil/magnetic bearing system designed by MiTi® for these tests has been tested with an applied load of almost 1350 lbs combined with a dynamic load of approximately 100 lbs. This bearing system is shown in Figure 7.101. Measured performance of the 100 mm diameter bearing is shown in Figure 7.102. As seen the combination of the two bearings exceeds the capacity of either alone.

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Principles of Gas Turbine Bearing Lubrication and Design

7-111

Foil brg

Foil brg

FIGURE 7.101 Hybrid foil magnetic journal bearing.

To demonstrate the ability of the foil bearing component of the hybrid bearing to act as a backup bearing for the magnetic bearing, the system response to potential magnetic bearing failures and failure recoveries was investigated. For this investigation, 13 different simulated failures were investigated at speeds of 15,000 rpm and 25,000 rpm. The failure modes selected for this investigation were: 1. 2. 3. 4. 5. 6. 7. 8.

Vertical Sensor Failure Horizontal Sensor Failure Both Sensors Failure Loss of Vertical Displacement Signal Loss of Horizontal Displacement Signal Loss of Both Displacement Signals Loss of Left Horizontal Amplifier Output Loss of Both Horizontal Amplifier Outputs

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology 1400 Hybrid foil/mag bearing

Permissible load capacity (lb)

1200 1000 800

Foil bearing

600 400 200

Magnetic bearing

0 0

FIGURE 7.102

5

10

15 20 Speed (krpm)

25

30

35

Experimentally measured performance of hybrid foil magnetic bearing.

Displacement (Mil)

3 2 Recovery

Failure

1 0 –1 –2 0

2

4

6

8

10

12

14

16

18

Time (sec)

FIGURE 7.103

Rotor response due to vertical sensor open failure at 25,000 rpm.

Displacement (mil)

3 Recovery

Failure

2 1 0 –1 –2 0

5

10 Time (sec)

FIGURE 7.104

Loss of vertical displacement signal at 15,000 rpm.

© 2006 by Taylor & Francis Group, LLC

15

Principles of Gas Turbine Bearing Lubrication and Design

9. 10. 11. 12. 13.

7-113

Loss of Top Vertical Amplifier Output Loss of Both Vertical Amplifier Outputs Loss of Bias Current Loss of All Control Amplifier Outputs Loss of All Amplifier Outputs and Bias Current

To simulate a worst case scenario, the bearing load was supported almost entirely by the magnetic bearing prior to the failure event. In addition, lightly damped magnetic bearing characteristics were selected for use during this testing. Based on previous experience, MiTi engineers examined both the initial failure transient as well as the system recovery In many cases, the recovery transient was actually more severe than the failure transient under these operating conditions. Figure 7.103 presents a typical shaft vertical response at the foil bearing displacement sensors for a failure that simulates sensor or cable damage which results on a loss of electrical continuity between the sensor and the sensor signal conditioning unit (-24 VDC control system input) and subsequent recovery of sensor signals. Horizontal shaft motion for this test was negligible. Figure 7.104 presents a typical shaft vertical response at the foil bearing displacement sensors for any system failure, which results in a loss of the vertical displacement measurement (0 VDC control system input). Again, the time-history trace of shaft displacement shows both failure and recovery. As was noted previously, horizontal shaft motion was negligible. In all 26 test cases, the transient during both failure and recovery was well controlled by the foil bearing. In addition, a foil bearing alone coast-down from full speed under shaft load was demonstrated in later testing. The combination of large static load and simulated failures demonstrate that the foil bearing component of a foil/magnetic hybrid bearing is an effective back up bearing for the magnetic bearing, allowing both continued operation following a failure, as well as damage-free equipment shut down and that the hybrid bearing system can support more load than either alone. 7.8.1.3 Compliant Foil Seals Due to the need to track rotor dynamic excursions without loss of seal performance either during the excursion or thereafter, advanced dynamic seals are needed. As with advances in bearings the improved low leakage film riding seals are needed to compliment the life and performance gains possible by the oil-free bearings. A Mohawk Innovative Technology, Inc. (MiTi®) novel, non-contacting, Compliant Foil gas Seal (CFS) has been designed and successfully tested at temperatures to 600◦ C and surface velocities of over 2100 ft/sec in a dynamic simulator representative of a small gas turbine engine hot section. Measured and analytical comparisons of leakage flow rates at varying differential pressures were made, showing that the CFS capability significantly exceeds the performance of both brush and labyrinth seals. The brush and CFS tests were performed with the rotor operating at speeds to 48,000 rpm. The labyrinth and CFS comparisons were made under nonrotating conditions, but with each seal mounted in the rig. Besides the performance benefits, the CFS offers improved life and durability benefits when applied to most rotating machinery, as noted later on. The performance of both the CFS, like the foil bearing, is based on the hydrodynamically generated high-pressure gas film that is built up between the journal and the seal top surface (top foil) due to the shaft rotation. This thin gas film separates the seal surface from the rotating shaft sealing surface, creating a film pressure (See Figure 7.105) resulting in noncontact, continuous operation. Strict accounting of all variables results in a seal with optimized performance attributes. While its primary purpose is the control of leakage, the similarity to a CFB results in its having a load-bearing capacity as a secondary function. This capacity is predictable and can be incorporated in the overall system design to provide additional support stiffness and damping for improved rotor system dynamics. Through compression system aerodynamic improvements, advanced materials, and weight reducing designs, improved engine efficiencies are being achieved. However this emphasis on efficiency, reduced emissions, and lowered noise levels is placing more stringent requirements on operating systems, including higher shaft surface speeds and increased environmental temperatures. This increased severity causes greater wear and mechanical distress for the seals due to their material and design limitations [47].

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Handbook of Lubrication and Tribology Compliant foil seal pressure profile including top foil

Pressure (Psi)

60 45 30 15 1

FIGURE 7.105

2 Ax 3 ial pos 4 5 itio n

100

6

50 75 on l directi ferentia Circum

25

CFS pressure profile. Compliant gas foil seal concept to application Shoulder

PH Top foil

Shaft

PL Spring bump Fabrication 1.4 in.

Concept

2.84 in.

FIGURE 7.106

5.950 in.

8.5 in.

CFS sizes.

An important conclusion from this discussion is that seals need improvements to accommodate the influence of operating speed, temperature, and pressure on seal performance and life. The higher operating speeds significantly affect seal heat generation and subsequent thermal distortions of mating components. Thus, as pressure–velocity product increases, seal durability, reliability, performance, and cost may all be adversely affected. Consequently, new materials, new designs, or some combination of both, are needed. Noncontacting labyrinth seals are insufficient due to the large gaps and hence leakage flows. Contact seals such as brush seals experience wear. Improved air seals will provide increased engine performance and improved life-cycle costs [47–49]. Advanced engine seals must have reduced losses and they must maintain these performance benefits over the service interval of the engine [50]. However, to achieve these benefits as surface speeds and ambient temperatures increase, new technologies that result in extremely lightweight design concepts, employ improved or novel sealing approaches with reduced leakage, and take advantage of new materials are needed. Damage-tolerant, high-performance designs are required

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Principles of Gas Turbine Bearing Lubrication and Design

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6 in. seal rotor motion under 80 psig differential pressure 90 120

mi

60

5 4

150

30

3 2 e 1

180

a

4

3

2

1

d

b c

0 5

f g

0

1

2

0 3

4

1

5

mi

Speed (rpm)

Temperature (°F)

a = 1500 b = 3000 c = 4000 d = 5000 e = 6000 f = 7000 g = 8000

72 73 76 75 75 74

2 3

210

330

4 5 300

240 270

FIGURE 7.107 Demonstrated CFS rotor excursion capability.

to accommodate the high differential pressures, extreme temperatures, high surface speeds, and normal rotor excursions. Advanced seals must also accommodate large rotor excursions; whether due to axial thermal gradients, maneuvers in aerospace applications, bowed rotor starts, and passage through bending critical speeds. Attempts to meet these stringent requirements with fixed geometries (i.e., labyrinth seals) has generally resulted in design compromises with larger than desired steady-state operating clearances. Brush seals have been investigated with a fairly high degree of success in numerous military and commercial engines and seem to address some of these limitations. For example, brush seals have shown leakage rates from 2 12 to 10 times lower than labyrinth seals and have demonstrated performance improvements in engine applications [50–52]. Although brush seals have demonstrated significantly lower leakage than conventional designs, contacting seals are subject to wear, which degrades their effectiveness with time. Good brush seal performance requires a smooth, wear resistant runner interface and wear resistant bristles. Further, the ability of brush seals to continue to operate reliably and efficiently under large rotor excursions is very limited. Experience has shown that the bristles plastically deform under large excursions, thereby changing the seal clearance and increasing leakage. As reported at the 1997 Seal Workshop at NASA/GRC [52], there are a number of concerns related to brush seals. Specifically, brush seals are costly than existing seals; they are heavier; the suppliers are limited; there is considerable concern over the durability; and there is concern over damage during construction and installation. In particular, care must be taken during installation to prevent reverse rotation of the shaft, which can damage the seal. While significant efforts are being expended to address the limitations present in brush seals, it will always be a contacting seal and thus prone to wear. In essence then, the performance of the brush seal is a function of time, shaft excursions, environment, and the tribo-material system. Given this background it is evident that there is a need for a highly efficient, long-life seal that is capable of maintaining performance during and after large excursions. Recent developments [53–56] indicate that noncontact CFB technology has great application and development potential for use as aircraft gas turbine engine seals. Dr. Heshmat of Mohawk Innovative

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Technology, Inc., (MiTi) has developed CFB’s with load capacities and damping significantly greater than previously demonstrated. This bearing technology foundation has been instrumental in the development of the CFS as reported by Salehi and Heshmat [57–59]. They have demonstrated the successful operation of a 1.4 , a 2.84, and a 6.0 in. CFS, and have even manufactured an 8.5 in. CFS for NASA that was tested to 30,000 rpm (See Figure 7.106). The current seal test has shown us that the lift off process and film riding does take place to produce the desired noncontact operation. The seal has also shown large excursion capability in excess of 6-mi (diametral) for a 150 mm diameter seal with an 80 psig differential pressure and while operating at speeds to 14,000 rpm at MiTi® (see Figure 7.107).

References The selection of the following references was made with the intent of providing sources from which additional data could be culled for bearing design purposes. Della Corte, C. and Pinkus, O., “Tribological Limitations in Gas Turbine Engines: A Workshop to Identify the Challenges and Set Future Directions,” ASME Publication Allaire, P.E., Li, D.F., and Choy, K.C., “Transient Unbalance Response of Four Multilobe Journal Bearings,” J. Lubr. Technol., Trans. ASME, July 1980. Chen, H.M. “Active Magnetic Bearing Technology: A Conventional Rotordynamic Approach,” Proceedings of 15th Leeds-Lyon Symposium on Tribology, Sept., 1988. Chen, H.M., “Magnetic Bearings and Flexible Rotor Dynamics,” Proceedings of STLE Annual Meeting at Cleveland, Ohio, May 9–12, 1988. Chen, H.M. et al., “Stability Analysis for Rotors Supported by Active Magnetic Bearings,” Proceedings of 2nd International Symposium on Magnet Bearings, July 12–14, 1990, Tokyo, pp. 325–328. Chen, H.M., “Design and Analysis of a Sensorless Magnetic Damper,” presented at ASME Turbo Expo, June 5–8, 1995, Houston, Texas, 95GT180. Heshmat, H. and Chen, H.M., Compressor handbook, Chapter 19, “Principles of Bearing Design,” McFrawHill, 2001. Gross, W.A., “Gas Film Lubrication,” John Wiley, 1962. Heshmat, H., Walowit, J.A. and Pinkus, O., “Analysis of Gas-Lubricated Compliant Thrust Bearings,” ASME Paper 82-LUB-39, 1982. Heshmat, H., Walowit, J.A., and Pinkus, O., “Analysis of Gas-Lubricated Foil Journal Bearings,” ASME Paper 82-LUB-40, 1982. Heshmat, H., and Dill, J. “Fundamental Issue in Cryogenic Hydrodynamic Lubrication,” Proceedings of AFOSR/ML Fundamentals of Tribology Work Shop, (February 1987). Heshmat, H. “Analysis of Compliant Foil Bearings with Spatially Variable Stiffness,” presented at AIAA/SAE/ASME/ASEE 27th Joint Propulsion Conference, June 24–26, 1991, Sacramento, CA, Paper No. AIAA-91-2101. Heshmat, H. “A Feasibility Study on the Use of Foil Bearings in Cryogenic Turbopumps,” presented at AIAA/SAE/ASME/ASEE 27th Joint Propulsion Conference, June 24–26, 1991, Sacramento, CA, Paper No. AIAA-91-2103. Heshmat, H. and Hermel, P., “Compliant Foil Bearing Technology and Their Application to High Speed Turbomachinery,” Proceedings of 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, Sep. 1993, D. Dowson et al. (Eds.), Elsevier Science Publishers B.V., 1993, pp. 559–575. Heshmat, H. and Pinkus, O. “Performance of Starved Journal Bearings with Oil Ring Lubrication,” J. Trib. Trans. ASME, 107, 1985, 23–32. Heshmat, H. and Pinkus, O. “Experimental Study of Stable High-Speed Oil Rings,” J. Trib. Trans. ASME, 107, 1985, 14–22.

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Heshmat, H. and Pinkus, O. “Performance of Oil Ring Bearing,” International Science Conference on Friction, Wear, Lubr., Tashkent, May 1985. Hustek, J.F. and Peer, O.J. “Design Considerations for Compressors with Magnetic Bearings,” Proceedings of 3rd International Symposium on Magnetic Bearings, July 93, Alexandria, VA. Jones, G.J. and Martin, F.A. “Geometry Effects in Tilting-Pad Journal Bearings,” ASLE Paper No. 78-AMatA-2, 1978. Ku, C.-P.R. and Heshmat, H. “Compliant Foil Bearing Structural Stiffness Analysis: Part I–Theoretical Model Including Strip and Variable Bump Foil Geometry,” J. Trib. Trans. ASME, 114, 1992, 394 –400. Pinckney, F.D. and Keesee, J.M., “Magnetic Bearing Design and Control Optimization for a Four-Stage Centrifugal Compressor,” Proceedings of Mag. ’92, pp. 218–227. Pinkus, O. and Sternlicht, B., “Theory of Hydrodynamic Lubrication,” McGraw-Hill, 1961. Pinkus, O. and Wilcock, D.F., “Low Power Loss Bearings for Electric Utilities: Volume II: Conceptual Design and Optimization of High Stability Journal Bearings; Volume III: Performance Tables and Design Guidelines for Thrust and Journal Bearings,” MTI Report Nos. 82TR42, 82TR43, April 1982. Pinkus, O., “Analysis of Elliptical Bearings,” Trans. ASME., Vol. 78, 1956, pp. 965–973. Pinkus, O., “Analysis and Characteristics of the Three-Lobe Bearing,” Trans. ASME, Series D., 81, 1959. Pinkus, O., “Solution of the Tapered-Land Sector Thrust Bearing,” Trans. ASME, 80, 1958. Pinkus, O., “Analysis of Non-circular Gas Journal Bearings,” J. Lubr. Technol., Trans. ASME, Oct. 1975. Pinkus, O. “Solution of Reynolds Equation for Arbitrarily Loaded Journal Bearings,” Trans. ASME, Series D, 83, 1961. Pinkus, O. “Misalignment in Thrust Bearings Including Temperature and Cavitation Effects,” J. Tribol., Oct. 1986. Pinkus, O. “Optimization of Tilting Pad Journal Bearings Including Turbulence and Thermal Effects,” Isr. J. Technol., 22, 1984/85. Pinkus, O., “Manual of Bearing Failure and Repair in Power Plant Rotating Equipment,” EPRI, July 1991. Raimondi, A.A. and Boyd, J., “A Solution for the Finite Journal Bearing and Its Application to Analysis and Design — III,” Trans. ASLE, 1, 1959. Reddickoff, J.M. and Vohr, J.H., “Hydrostatic Bearings for Cryogenic Rocket Engine Turbopumps,” J. Lubr. Technol., 1969. Schmied, J.L. and Predetto, J.C., “Rotor Dynamic Behaviour of a High-Speed Oil-Free Motor Compressor with a Rigid Coupling Supported on Four Radial Magnetic Bearings,”proceedings of 4th International Symposium on Magnetic Bearings, August 23–26, 1994, ETH Zurish, Switzerland, pp. 441–447. Vohr, J.H., “The Design of Hydrostatic Bearings,” Columbia University, NY. Walton, J.F. and Heshmat, H.,“Compliant Foil Bearings For Use in Cryogenic Turbopumps,” Proceedings of Advanced Earth-to-Orbit Propulsion Technology Conference Held at NASA/MSFC May 17–19, 1994, NASA CP3282, Vol. 1, Sept. 19, 1994, pp. 372–381. Wilcock, D.F. and Booser, E.R., “Bearing Design and Application,” McGraw Hill, 1957. Suriano, F.J., Dayton, R.D., and Woessner, F.G., “Test Experience with Turbine-End Foil Bearing Equipped Gas Turbine Engines.” ASME Paper No. 83-GT-73. Heshmat, H., Chen, H.M., Walton, J.F., 1998, “On the Performance of Hybrid Foil-Magnetic Bearings,” Proceedings of 43rd ASME Gas Turbine and Aeroengine Congress, Stockholm, Sweden, ASME Paper No. 98-GT-376. Heshmat, H., “Advancement in the Performance of Aerodynamic Foil Journal Bearings- High Speed and Load Capacity,” ASME J. Tribol., 116, 1994, 287–95. Heshmat, H. and Ku, C.-P., “Structural Damping of Self-acting Compliant Foil Journal Bearings.” ASME J. Tribol., 116, 1994, 76–82. Heshmat, C.A. and Heshmat, H., “An Analysis of Gas-Lubricated, Multileaf Foil Bearings with Backing Springs.” ASME J. Tribol., 117, 1995, 437–443. Heshmat, H. and Hermel, P., “Compliant Foil Bearing Technology and their Application to High-speed Turbomachinery,” Proceedings of 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, D. Dowsen, et al., (Eds.), Elsiver, 1992, pp. 559–575.

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Walton, J.F. and Heshmat, H. “Compliant Foil Bearings for use in Cryogenic Turbopumps.” NASA CP3232 Vol. 1, 1994, pp. 372–381. Heshmat, H., “Operation of Foil Bearings Beyond the Bending Critical Mode.” ASME Paper 99-TRIB-48. Swanson, E.E., Walton, J.F., Heshmat, H., 1999, “A 35,000 RPM Test Rig for Magnetic, Hybrid, and Back-Up Bearings.” ASME Paper No. 99-GT-180, 1999. Swanson, E., Heshmat, H, and Shin, J., “The Role of High Performance Foil Bearings in an Advanced, Oil-Free, Integral Permanent Magnet Motor Driven, High-Speed Turbo-Compressor Operating above the First Bending Critical Speed,” Proceedings of Turboexpo 2002, ASME Turbo Expo 2002, June 3–6, Amsterdam, The Netherlands, 2002, GT-2002-30579. NASA. “Starting a Turbomachinery Revolution” NASA Glenn Research Center (216) 433-5573 Heshmat, H. “Advancements In The Performance of Aerodynamic Foil Journal Bearings: High Speed and Load Capability,” ASME Paper 93-Trib-32, STLE/ASME Tribology Conference, October 24–27, New Orleans, LA, 1993. Heshmat, H. “Advancements In The Performance of Aerodynamic Foil Journal Bearings: High Speed and Load Capability,” ASME Paper 93-Trib-32, STLE/ASME Tribology Conference, October 24–27, New Orleans, LA, 1993. Hendricks, R.C., Griffin, T.A., Kline, T.R., Csavina, K.R., Pancholi, A., and Sood, D., “Relative Performance Comparison Between Baseline Labyrinth and Dual-Brush Compressor Discharge Seals in a T700 Engine Test,” proceedings of 39th ASME International Gas Turbine and Aeroengine Conference, The Hague, Netherlands, June 13–16, 1994. Paper No. 94-GT-266 Flower R., “Brush Seal Development Systems, AIAA Paper 90-2143, 1990. Hendricks, R. and Steinitz, B. Editors, “Seals/Secondary Flows Workshop 1997, V.1, October 1998. Heshmat, H. and Hermel, P. “Compliant Foil Bearing Technology and Their Application to High Speed Turbomachinery,” proceedings of the 19th Leeds-Lyon Symposium on Thin Film in Tribology — From Micro Meters to Nano Meters, Leeds, Sep. 1992. Heshmat, H. “Analysis of Compliant Foil Bearings with Spatially Variable Stiffness,” proceedings of AIAA/SAE/ ASME/ASEE 27th Joint Propulsion Conference, June,1991, Sacramento, CA, Paper AIAA-91-2102. Heshmat, C.A. and Heshmat H. “An Analysis of Gas Lubricated, Multi-Leaf Foil Journal Bearings with Backing Springs,” ASME Paper 94-Trib-61. Heshmat, H. and Ku, C-P R., “Structural Damping of Self-Acting Compliant Foil Journal Bearings,” ASME Trans., J. Tribol., 116, 1994, pp. 76–82. Salehi, M., Heshmat, H., Walton, J., and Cruzen S., “The Application of Foil Seals to a Gas Turbine Engine,” AIAA paper 99-2821, proceedings of 35th AIAA/ASME/SAE/ASEE joint Propulsion Conference and Exhibit, June 20–24, Los Angeles, CA, 1990. Salehi, M, and Heshmat, H., 2000, “High Temperature Performance Evaluation of a Compliant Foil Seal,” presented at 36th AIAA/ASME/SAE/ASEE joint Propulsion Conference and Exhibit, July 17–19, Huntsville, AL, 2000. Salehi, M., and Heshmat, H., “On the flow and thermal analysis of the compliant gas foil seals and foil bearings,” STLE Trans., 43, 2000, pp. 318–324.

Additional References Salehi, M., Heshmat, H., and Walton, J.F., II “On the Frictional Damping Characterization of Compliant Bump Foils,” Presented at the International ASME/STLE Joint Tribology Conference, October 2002, Cancun, Mexico, ASME J. Tribol., 125, 2003, pp. 804–813. Salehi, M., Heshmat, H., and Walton, J.F., II “Advancements in the Structural Stiffness and Damping of a Large Compliant Foil Journal Bearing — An Experimental Study,” Paper GT2004-53860, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004.

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Salehi, M., Heshmat, H., Walton, J.F., II and Tomaszewski, M.J. “Operation of a Mesoscopic Gas Turbine Simulator at Speeds in Excess of 700,000 rpm of Foil Bearings.” Paper GT2004-53870, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004. Walton, J.F., II, Heshmat, H., and Tomaszewski, M.J. “Testing of a Small Turbocharger/Turbojet Sized Simulator Rotor Supported on Foil Bearings.” Paper GT2004-53647, Published in the ASME J. Eng. Gas Turbines Power, and presented at the International Gas Turbine Institute Conference, Vienna, Austria, June 2004. Hryniewicz, P., Locke, D.H. and Heshmat, H. “New-Generation Development Rigs for Testing High-Speed Air-Lubricated Thrust Bearings.” Tribol. Trans., 46, 2003, pp. 556–569. Heshmat, C.A., Xu, D.S., and Heshmat, H. ”Development of Advanced Thrust Foil Bearings.” Presented at the 54th STLE Annual Meeting, May 23–27, Las Vegas, NV, 1999. Submitted for publication in the Tribology Transactions. Heshmat, C.A., Xu, D., and Heshmat, H. “Analysis of Gas Lubricated Foil Thrust Bearings using Coupled Finite Element and Finite Difference Methods.” Presented at the ASME/STLE Joint Tribology Conference, October 10–13, Orlando, FL, 1999, published in the J. Trib. Trans. ASME, 122, 2000, pp. 199–204. Heshmat, C.A., Xu, D.S. and Heshmat, H. “Analysis of Gas Lubricated Foil Thrust Bearings Using Coupled Finite Element and Finite Difference Methods,” Trans. ASME, J. Tribol., 122, 2000, pp. 199–204.

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8 Steam Turbines 8.1 8.2

Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Features of Steam Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-1 8-2

Classification

8.3

Turbine Design and Construction . . . . . . . . . . . . . . . . . . . .

8-3

Bearings • Bearing Housings and Bearing Housing End Seals • Steam Control Valves, Governors, and Control Systems • Turning Gear • Couplings • Additional Tribological Components and Issues • Driven Units

8.4

Lube Oil Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-17

Nonpressurized Oil Ring Lubrication • Pressurized Lubrication Systems

8.5

Turbine Oil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-23

Physical Properties • Formulation

8.6

Performance Features of Turbine Oils. . . . . . . . . . . . . . . .

8-26

Viscosity • Oxidation Stability • Freedom from Sludge and Deposits • Corrosion Protection • Water Separability (Demulsibility) • Air Separability and Resistance to Foaming

8.7

Degradation of Turbine Oils in Service . . . . . . . . . . . . . .

8-28

Contamination • Additive Depletion • Thermal and Oxidative Degradation • Biological Deterioration • Turbine Oil Severity

8.8

Lubricant Maintenance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-30

New Oil Makeup • Lube Oil Purification • Refortification

8.9

B.C. Pettinato Elliott Company

Fire-Resistant Fluids . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-32

Properties • Degradation • Condition Monitoring • Maintenance

References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-34

8.1 Introduction Steam turbines are used extensively in the power generation industry as prime movers for generators. They are also used for mechanical drive application in petrochemical and other industries where they power centrifugal pumps, compressors, blowers, and other machines. In addition, they continue to be used for shipboard propulsion. Sizes range from as low as 0.75 kW for some mechanical drive applications to as high as 1,500 MW for electric generator drives in large nuclear power plants [1]. Steam turbines are particularly well suited for continuous operation, and in many cases are operated for years without shutting down.

8-1

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Handbook of Lubrication and Tribology

8.2 Features of Steam Turbines Steam turbines operate by taking high-pressure steam and converting it into useful mechanical work through expansion. The steam is fed into an inlet casing then throttled through a set of inlet valves, which control the rate of steam admission into the turbine. The steam is then allowed to expand and accelerate through stationary blades or nozzles, which directs the flow onto the rotating blades. The rotating blades convert the steam’s kinetic energy into torque, which results in rotation of the turbine shaft along with a loss of pressure and temperature in the steam. The rotating shaft is used to drive machinery coupled to the exhaust end of the turbine shaft. Absence of lubrication from the steam path is an important feature. Since the exhaust steam is not contaminated with oil vapor, this allows the steam to be condensed and returned directly to the boilers for reheat, or extracted and used for direct heating or other purposes. The lack of internal lubrication also results in a relatively low rate of lubricating oil consumption [2].

8.2.1 Classification Steam turbines have numerous configurations and means of classification. A steam turbine is generally classified as being either high-pressure or low-pressure, condensing or noncondensing, single-stage or multi-stage, single-valve or multi-valve, extraction or nonextraction, direct drive or gear drive, and for either electric generator, mechanical drive, or propulsion service [3]. In addition, steam turbines are classified in accordance with recognized engineering standards, which govern various aspects of turbine design and construction. Some of these classifications are discussed further. High-pressure designs refer to the internal pressure to be contained by the main shell and casing parts. High pressure generally refers to pressures in excess of 13,800 kPa (2,000 psig) where double shell construction is often used. The pressure and temperature of steam are interrelated. Higher inlet steam pressure is often accompanied by higher steam temperature. Temperatures can range from 200◦ C to over 600◦ C. High temperature generally refers to applications with inlet temperatures in excess of 540◦ C (1,000◦ F). Condensing turbines exhaust steam at less than atmospheric pressure, whereas noncondensing (back pressure) turbines exhaust steam at higher than atmospheric pressure. Condensing machines tend to be larger and more complex than noncondensing designs due to the increased volume expansion of the steam at the exhaust end as well as the additional hardware required to drop the exhaust end pressure below atmospheric. In direct drive arrangements, the turbine is directly coupled to the driven machine; whereas gear drive applications have either a speed increasing or speed reducing gear between the turbine and driven equipment. The use of a speed increasing or reducing gear creates added complexity, cost, and power losses along with additional requirements of the lube oil system. However, the use of gears greatly increases the application range whether the need is for high torque as in marine propulsion or high-speed requirements such as integrally geared compressors. Gear drives also enable the efficient use of small turbines, which can operate at higher speeds when a reduction gear is used. Generator drive turbines operate at single speeds to synchronize the generators with the electric grid. Typically, the synchronization speed is either 1,800 or 3,600 rpm in regions with 60 Hz power, or 1,500 or 3,000 rpm in regions with 50 Hz power. On the other hand, mechanical drive turbines are variable speed with shaft speeds as low as 1,000 rpm or as high as 20,000 rpm depending on the turbine and the application. A number of different engineering standards have been developed for the design and procurement of steam turbines as shown in Table 8.1. American Petroleum Institute (API) standards pertain to design, manufacture, and testing of mechanical drive turbines for petrochemical application [4,5]. National Electrical Manufacturers Association (NEMA) standards pertain to design and application of mechanical drive turbines and turbine generator sets for electric utility application [6,7]. Military standards generally apply to steam turbines for shipboard use [8–10]. Other international

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Steam Turbines TABLE 8.1

8-3 Steam Turbine Design and Procurement Standards

Standard designation API 611 API 612 / ISO 10437 IEC 60045-1 NEMA SM-23 NEMA SM-24 MIL-T-17286D MIL-T-17600D MIL-T-17523

Standard title General-purpose steam turbines for refinery service Special-purpose steam turbines for refinery service Steam turbines — part 1: specifications Steam turbines for mechanical drive service Land-based steam turbine generator sets 0 to 33,000 kW Turbines and gears, shipboard propulsion, and auxiliary steam; packaging of Turbines, steam, propulsion naval shipboard Turbine, steam, auxiliary (and reduction gear) mechanical drive

recognized standards such as IEC 60045-1 are also used to assist in steam turbine specification and procurement [11]. Figure 8.1 shows a general-purpose (API 611) turbine. These turbines are either horizontal or vertical units used to drive equipment that is usually spared, is relatively small in size (power), or is in noncritical service. General-purpose steam turbines for refinery service are intended for applications where the inlet gauge pressure does not exceed 4,800 kPa (700 psi), the inlet temperature does not exceed 400◦ C (750◦ F), and the speed does not exceed 6,000 rpm [4]. The turbine shown in Figure 8.1 has lubrication consisting of sumps at each journal bearing with oil ring-lubricated bearings. An isolated mechanical–hydraulic governor with oil sump is used to control speed. Figure 8.2 shows a special-purpose turbine for refinery application that meets API 612/ISO 10437 specifications. Such units are usually not spared and are used in uninterrupted continuous operation in critical service. They are not limited by steam conditions, power, or turbine speed. The equipment (including auxiliaries) covered by these standards are designed and constructed for a minimum service life of 20 yr and at least 5 yr of uninterrupted operation [5]. The turbine shown in Figure 8.2 has lubrication provided by a circulating oil system console (not shown) providing oil at high volumes to the bearings and to the servo valve actuator.

8.3 Turbine Design and Construction The parts of a steam turbine may be thought of as being in four groupings (1) the rotor, or spindle, (2) stationary parts, (3) the governing and trip systems and valves, and (4) auxiliary systems consisting of the lubrication system and other components such as the condition monitoring system. The rotor, depending on turbine type, may consist of wheels mounted on a shaft or may be machined from a solid forging or a forging made up of welded sections. In each case, the rotor carries securely fastened radial blades or buckets. Principle stationary parts consist of the steam-tight casing, nozzles, shaft seals, and bearings. Turbine governors control speed by controlling steam-admission valves through mechanical, pneumatic, or hydraulic actuators. Those parts of the turbine requiring lubrication consist of the bearings supporting the rotor, hydraulic actuators and governor components, and the trip system; and in some cases, a turning gear, geared couplings, and front standard. Lubricated parts reside external to the steam path, and when properly isolated will not contaminate the steam or become contaminated by the external environment. The lubrication system may be simple reservoirs in the pedestals of ring-oiled bearings, or elaborate circulation systems, having pumps, coolers, filters, and monitoring devices [12]. Figure 8.3 shows a typical unit of an oil-piping diagram for a turbine, gear, and generator string. Lubricating oil is supplied at two pressures by an oil console (not shown). Lube oil is supplied at low pressure of 100 to 125 kPa (15 to 18 psig) to the bearings. High-pressure oil of 1,000 kPa (150 psig) is supplied to the trip and throttle valve, to the valve actuator, and, if needed, to the governor mechanism. Bearing and coupling housings are part of the lube oil circuit and act to return oil to the reservoir.

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8-4

Sentinel valve

Rotor disk assembly

Casing cover

Shaft sleeve seal

Exhaust end sealing gland Carbon ring assembly

Steam end sealing gland

Rotor locating bearing Overspeed thip assembly Coupling (governor drive)

Carbon ring assembly

Governor Governor linkage

Oil rings

Oil rings

Rotor shaft

Steam end bearing housing

Shaft sleeve seal Steam chest Governor valve

Exhaust end bearing pedestal Exhaust end casing

FIGURE 8.1

Reversing blade assembly

Steam end casing Nozzle ring

Steam end journal bearing

Steam end support

General-purpose steam turbine. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.)

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Handbook of Lubrication and Tribology

Exhaust end journal bearing

Steam Turbines

Rocker arm bearing

Lubrican connection

Governor linkage assembly Valve stem and packing Valves, seat, and bar assembly Breather cap Bearing Journal housing bearing Bearing housing deflector Shaft end with coupling bolt pattern

Steam chest

Turbine case

Gland packing assembly case

Exhaust end packing gland assembly

Breather cap Interstage shaft seals

Journal bearing Thrust bearings

Steam end packing gland assembly

Rotor

Bearing housing end seal

Oil drain Steam exhaust

Bearing housing end seals

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Gland seal leak off

8-5

Steam turbine for special-purpose refinery service.

Bearing housing

Steam end flexible support Casing drain

FIGURE 8.2

Oil drain

8-6

Lube oil supply to unit conn’s

P1

Flexible hose

PSH PSLL

S P1

Breather Servo motor Driven equipment

IP

Coupling

Breather Gear

To reservoir

FIGURE 8.3

SG

SG

1/2” per foot minimum

Concentric reducer I

Inlet SY servo motor/ valves

Trip and throttle valve S

e

Slop

Thermowell Thermowell

P

Current/pneumatic convertor

PCV

Pressure control valve

P1

Pressure indicator

PSH

Pressure switch, high

PSLL

Pressure switch, trip

SG

Site glass

SY

Speed relay

T1

Temperature indicator

T1

T1

SG

SG

Unit oil piping diagram for turbine-gear-generator set.

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Orifice PSLL

SG

Handbook of Lubrication and Tribology

SG

T1

2 valve manifold instrument valve/bleed valve

High pressure control oil supply

Trip solenoid dump valve

Spare vent

Spare vent

Solenoid valve (2-way) Ball valve

Nitrogen precharge

Turbine

Coupling

SG

Governor control signal

N2

Control oil accumulator

Steam Turbines

FIGURE 8.4

8-7

Tilt pad journal bearing. (From Elliott Company. With Permission.)

8.3.1 Bearings Proper rotor position is maintained by journal and thrust bearings. Journal bearings are used in pairs for radial positioning of the turbine shaft supporting the gravitational load of the shaft. Thrust bearings are used for axial positioning and support thrust loads that arise from steam forces within the turbine case. Thrust bearings are located at the steam end of the turbine opposite the coupling, and are used in pairs to accept thrust loading in either direction along the axis of the rotor. Steam turbine bearings can be either hydrodynamic, rolling element, or magnetic. Hydrodynamic bearings are the most prevalent. 8.3.1.1 Hydrodynamic Bearings Hydrodynamic bearings are highly advantageous because they suffer little or no wear and have exceptionally long life thereby enabling long periods of continuous operation, often in excess of 5 yr. In addition, the bearings possess dynamic characteristics that allow for vibration control thereby enabling high-speed operation, and traverse of rotor critical speeds. For this reason, hydrodynamic bearings are the most common type of bearing applied to steam turbines. Journal bearings are most often of the plain cylindrical, elliptical, multilobed, pressure dam, or tilt pad design. Figure 8.4 shows a schematic of a tilt pad journal bearing. Tilt pad journal bearing designs consist of several pads arranged in a ring around the shaft with the pads free to tilt about their respective pivots. Tilt pad journal bearings may include several design variations such as self-aligning features to compensate for misalignment, and special oil feed and drain configurations for temperature and power loss control [13,14]. One particular advantage of tilt pad journal bearings is their dynamic characteristics and inherent resistance to rotordynamic instability, which allows for control of vibration even at high speeds. Thrust bearings are usually of the tapered land or tilt pad design. Figure 8.5 shows a six shoe selfequalizing tilt pad thrust bearing. Tilt pad thrust bearings may also have features to compensate for misalignment, as well as special oil feed and drain configurations for temperature and power loss control [15,16]. Hydrodynamic bearings are lubricated with turbine grade oil either by a low-pressure circulating supply system or by ring lubrication where appropriate. In low-pressure supply systems, the oil flow is metered to each bearing by an orifice or other flow-controlling device. The oil flows into the clearance spacing of the bearing where it forms a wedge separating the bearing and shaft surfaces. The oil exits axially out the sides of journal bearings; and exits radially and tangentially from thrust bearings. Observation of drain oil flow through sight boxes can be taken as an indication of at least partial flow through the bearing and is often used as a quick indication that the oil pump is running and that the oil supply is probably sufficient. Oil supplied to the bearings functions as both a lubricant and as a coolant to

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8-8

Handbook of Lubrication and Tribology

FIGURE 8.5 Self-equalizing tilt pad thrust bearing. (From A General Guide to the Principles, Operation and Troubleshooting of Hydronamic Bearings, Publication HB, Kingsbury, Inc., Philadelphia, PA, 1997. With Permission.)

counteract the heat generated by shearing of the oil during operation and conduction from the hot rotor. Hydrodynamic bearings are limited with respect to minimum film thickness, maximum bearing temperature, and peak oil film pressure. These restrictions are inherently related to the load, speed of operation, and design of the bearing [17]. The bearings may be boundary lubricated during startup and turning gear operation, developing a full film shortly after startup. Operational film thickness is typically 25 to 75 µm (0.001 to 0.003 in.). Bearing metal temperature at the instrumented location may range from less than 55◦ C (130◦ F) for an unloaded inactive thrust bearing up to 130◦ C (265◦ F) for a bearing operating near its design limits. Peak oil film pressure is typically 2.5 to 3 times the specific load defined as P=

W A

(8.1)

such that P is the specific load (N/mm2 ), W is the load (N), and A is the projected area (mm2 ) [17]. For journal bearings, the area is the product of the diameter and length. For thrust bearings, the area is the area of the loaded surface. Bearing surfaces consist of a soft metal bonded to a hard metal backing. For North American operation, the soft metal surface is most often an ASTM B23 grade 2 babbitt comprised of 89% tin alloyed with antimony, lead, copper, iron, and trace amounts of other metals. Equivalent specifications can be found in ISO 4381 as SnSb8Cu4 [18], and Federal Spec Q-T-390 Grade 2. In some cases, an ASTM B23 grade 3 babbitt is used. Babbitt bearing surfaces generally cause the least damage to steel shafts when operated with inadequate lubrication or with contaminants. Babbitts are good for embedding hard contaminant particles and for resistance to seizure and galling [19]. In addition, tin-based babbitt is highly resistant to corrosion from organic acids and can provide satisfactory operation in the presence of oxidized and contaminated oils. A disadvantage of babbitt bearing materials is their relatively low compressive, tensile, and fatigue strengths especially at high temperature. To provide additional strength, the babbitt surface is cast and bonded as a thin layer to a hard metal backing, which may be steel, bronze, or chromium copper. Steel is the most prevalent and least expensive backing material. Chromium copper is used for its superior thermal conductivity enabling reduced bearing metal temperature. A good babbitt bond is critical, and can be inspected by nondestructive ultrasonic testing as described in ISO 4386 Part 1 [20]. The journal or thrust collar/disk is usually polished steel with surface finishes not exceeding 0.8 µm (32 µin.) Ra . The rotating element is either an integral part of the turbine shaft or else attached mechanically to the shaft. Bearing surface materials are normally steel containing less than 2.5% Cr, to prevent

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Steam Turbines

8-9

a type of failure known as wire wool [20]. In those cases where 12-Cr shaft material is required due to erosion concerns, a sleeve may be used as the bearing journal or the rotor may be inlaid with an acceptable steel material. To supervise the satisfactory and safe operation of turbine bearings, one or more of the following quantities may be monitored: inlet and drain temperature of oil, bearing metal temperature, drain flow of oil, shaft position, vibration amplitude, oil film pressure, and lift oil pressure. Bearing metal temperatures are measured using temperature sensors embedded in the bearing backing metal near the babbitt bond line and bearing surface [21]. High bearing metal temperature can be indicative of potential bearing failure. Bearing metal temperature that rises in an upward trend without corresponding change to load or speed is also indicative of potential bearing failure. There is varying opinion with respect to metal temperature limitation. In general, the manufacturer’s recommendation should be followed especially for new equipment lacking in historical data. Drain temperatures are also useful for identifying problems. Drain temperature of oil is an indicator of bearing power loss if measured separately for each bearing. It is also an indirect indication of bearing health, but not as reliable as bearing metal temperature. For this reason, drain oil temperature is not relied upon as an indication of safe operation unless the bearings are not instrumented. Radial and axial shaft position and vibration are measured with noncontacting eddy-current probes. Drastic shaft movement is an indication of bearing distress that occurs with wiping of babbitted surfaces [22]. Excessively high radial vibration is another sign of potential bearing distress and needs to be monitored as it may cause babbitt surfaces to fatigue or internal rubs to occur. Depending on the bearings, hydrodynamic bearing maintenance can consist of inspection, repair, or replacement. After installation, a lift check should be performed on each journal bearing, and the thrust bearing endplay and rotor axial position should be checked and recorded. These same checks should also be performed prior to bearing removal especially if abnormal bearing conditions were observed prior to shutdown such as high metal temperature or vibration. During visual inspection, the bearings are examined for signs of wear or distress such as scoring, cracks, pivot fretting or brinelling, heat discoloration, electrostatic discharge machining, corrosion, flaking, signs of overheated or contaminated oil such as varnish deposits, and loss of babbitt bond. The rotor journal and thrust areas also need to be examined for signs of distress such as scoring. Causes of bearing distress and failure include overloading, insufficient oil flow, insufficient bearing clearance or endplay, excessive overspeed, excessive vibration, and too high inlet oil temperature. Corrosion failures for tin babbitt bearings are fairly uncommon, but can occur in certain cases. The formation of hard deposits of tin oxide on tin rich white metal has been a problem with bearings in steam turbines caused by electrolytic action in certain environments such as when the oil contains free water with salt in solution [20]. Oil contamination from process gases that originate from the seal oil systems of driven units such as compressors can be particularly corrosive and may attack the components found in babbitt. 8.3.1.1.1 Hydrostatic Jacking Hydrodynamic bearings may include additional features such as an externally pressurized hydrostatic jacking system. The purpose of hydrostatic jacking is first to reduce the required breakaway torque during either start-up or turning gear operation and second to reduce bearing wear during turning gear operation. Hydrostatic jacking is effective by simply reducing the loading on the bearing surface such that it is within acceptable ranges. One manufacturer recommends consideration of hydrostatic jacking when the specific load on startup exceeds 1,300 kPa (190 psi) for plain journal bearings, 1,200 kPa (175 psi) for tilt pad journal bearings, and 60% of the maximum load for thrust bearings [23–25]. The need for hydrostatic jacking depends on the frequency of start-ups, duration of any baring condition, and available starting torque. Hydrostatic jacking systems are typically designed to lift the rotor off the bearings; however, this is not always practical. Hydrostatic jacking is effective so long as the friction torque is acceptable, the loading on the babbitt surface is reduced, and associated wear is negligible. Bearings with hydrostatic lift features require a high-pressure oil system, which typically supplies oil

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology

at 7,000 to 14,000 kPa (1,000 to 2,000 psig). The high-pressure oil is turned off shortly after start-up and turned on during coast down. Lift oil pressure may be indicated by reading the pressure supplied to the lift pocket. 8.3.1.2 Rolling Element Bearings Rolling element bearings (also known as antifriction bearings) are used where service is not critical or the steam turbine is spared. These bearings can be used as a complete set to accommodate both radial and axial loads or are used as a thrust bearing in conjunction with ring-lubricated bearings. Rolling element bearings are generally less reliable than pressure fed hydrodynamic bearings and are only applied when they meet specific criteria with respect to their speed and life, which are designated by dN and L10 parameters, respectively. The dN parameter is the product of d, the journal diameter (mm) and N , the rated speed in revolutions per minute. Operation of dN in excess of 300,000 generally requires oil lubrication. The L10 parameter describes the basic rating life expressed in number of operating hours, or millions of revolutions with 90% reliability. The latest revised and updated L10 equation considers the bearing design, dynamic load, reliability factor, and life adjustment factor that involves the complex interaction of lubrication conditions, contamination, bearing material properties and other factors [26]. Rolling element bearings are generally designed and retained in accordance with American Bearing Manufacturers Association (ABMA) standards. The Conrad type or deep groove ball bearing is a typical design. The bearings are lubricated either by grease with protection against overgreasing, or by oil supplied by bath, mist, or jet lubrication. Grease fittings are required to extend outside the machine to permit regreasing during operation. Venting is provided to prevent pressure buildup in the housing. One particular disadvantage of rolling element bearings is that they cannot be horizontally split without reducing their life and degrading their performance. As a result, most rolling element bearings cannot be replaced without removing the rotor and coupling. Presence of water in oil is particularly detrimental to the life of a rolling element bearing [27].

8.3.2 Bearing Housings and Bearing Housing End Seals Bearing housings support and position the bearings such that the rotor is centered in its respective packing bores. These housings are also used to mount vibration monitoring and other condition monitoring devices. The steam end bearing housing further encases the overspeed trip assembly; as well as the governor speed sensor, which may consist of a notched wheel and speed pickup, or it may consist of flyweights or other devices. At times, a turning gear is also present. Grounding brushes may be mounted to the outboard end of the bearing housing to prevent the buildup of high voltage between the shaft and the case, which can damage the bearings through electrostatic discharge. Bearing housings also function as a part of the lube oil circuit, keeping oil in while keeping contaminants, such as steam out. In the case of pressure-lubricated hydrodynamic bearings, the housings are arranged to minimize foaming through proper design of the drain and vent system to maintain oil and foam levels below shaft end seals. Proper sizing of drains is important to minimize foaming. Bearing housings are equipped with replaceable labyrinth end seals and deflectors where the shaft passes through the housing to minimize contamination and leakage. Bearing housings and gland seals are spaced to help prevent leaking oil from entering the glands and gland steam from entering the bearings. For ring-oiled bearings, the housings further act as oil sumps and may contain water jackets for cooling the oil. Bearings housing oil seals may suffer from oil carburization, contaminant leakage into the seal or oil leakage from the seal. Contaminant leakage into the bearing housing can be a problem when using a vapor extractor on the main oil tank, which creates a slight vacuum in the bearing housings through the oil drain lines. Pressurizing the annulus in the oil seal with a gas purge such as nitrogen or air can assist with

© 2006 by Taylor & Francis Group, LLC

Steam Turbines

8-11

seal leakage. This may also cool the oil seal to prevent carburization. Overheating of the oil seal may also be prevented by improvements to the heat guards [28].

8.3.3 Steam Control Valves, Governors, and Control Systems As shown in Figure 8.6, steam is directed into the turbine steam chest through either a trip or trip and throttle (T&T) valve. Trip valves are opened by fluid pressure and mechanically closed by spring force. The trip system is controlled by an overspeed governor. A trip due to overspeed or other unsafe operating condition causes the solenoid valve to open thereby causing system depressurization and immediate closure of the trip valve, which shuts off the steam thereby bringing the turbine to an eventual stop. After passing through the trip valve, the steam is directed through the steam chest, and then through control valves (also called governor valves). The control valves throttle the steam into a nozzle ring matching the turbine power to the load thereby controlling speed. The control valves may be operated by mechanical linkage, by bar-lift arrangement (Figure 8.6), by cams, or by individual hydraulic cylinders. Mechanical and pneumatic actuators can be found on the smallest turbines whereas hydraulic actuators are required on most other units. Extraction turbines have additional valves located at an intermediate stage in the turbine. Extraction valves may be of poppet or spool type for higher pressure, or of grid type for controlling large volumes of steam at lower pressure. In each case, the valve actuators are controlled by the main governor. The main governor operates independently from the overspeed governor. The main governor can be either a relatively simple system that acts directly upon a steam-admission valve; or a complex system that may control speed, extracted steam, and devices separate from the turbine such as a compressor or the boiler. Figure 8.7 shows a mechanical–hydraulic governor with hydraulic actuator. In this case, hydraulic accumulators are used to supply the high volume of fluid that is required for rapid control action during sudden changes in load. To achieve the high force levels required in multivalve applications, the governor typically controls a servo (prepilot or slave) to a master pilot that controls the flow of high pressure oil to a large piston as shown in Figure 8.7. The assembly of servo, pilot valve, and piston is called a servomotor. In such a control system, a few ounces in governor force can be multiplied through a hydraulic mechanical advantage to generate the thousands of pounds of force that may be required to operate the turbine governor valves [29]. Required hydraulic oil pressures typically range from 350 kPa (50 psi) on small turbines to 18,000 kPa (2,600 psi) on very large turbines [30,31]. Turbine oils are typically employed at pressures below 2,000 kPa (290 psi) whereas fire-resistant fluids are often used at pressures exceeding 2,000 kPa and in installations where steam pipe temperatures exceed the auto-ignition temperature of turbine oil, particularly in power plant applications [30]. The governor and actuator control system may be supplied from the same lube system as the bearings or may be fed independently from a separate system. In small turbines, hydraulic and mechanical–hydraulic governors are often self-contained units featuring a shaft driven oil pump, and an oil sump with sight glass for determining the oil level. In medium sized turbines, typical of process industry applications, the control system is normally fed off the same lube system as the bearings. In large power plant turbines, two separate circulating systems are usually employed: one for the bearings using turbine oil and one for the control system using a fire-resistant phosphate ester fluid. Control systems have long had high visibility due to reliability and maintenance shortcomings. Large quantities of mechanical components such as pins, links, levers, rod end bearings, hydraulic relays, springs, gearing, and flyball governor assemblies are present and subject to wear. The use of electronic speed sensors and electronic governor controls has enabled the elimination of some wearing mechanical parts, and has improved control and flexibility through use of noncontacting pick-ups and nonmechanical feedback circuits. Actuators have remained primarily hydraulic due to the large forces and quick response time required. Governor maintenance depends considerably on the type of governor in use, and the manufacturer’s recommendations should be followed. The proper oil must be selected, and it must be kept clean, dry, and at

© 2006 by Taylor & Francis Group, LLC

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High-pressure oil supply from oil console Variable-pressure control oil Drain oil to oil console reservoir

Steam to turbine

Trip and throttle valve

Inlet steam valves

Trip pin

Spring-loaded handle Solenoid valve

Trip lever Knife-edge

Bearing housing Servo motor

Orifice

High-pressure oil from oil console

FIGURE 8.6

To oil console drain

Trip system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)

© 2006 by Taylor & Francis Group, LLC

Handbook of Lubrication and Tribology

Electrical leads

Steam chest

Variable-pressure control oil

Steam Turbines

KEY High-pressure oil supply from oil console

Drain oil to oil console reservoir Governor internal high-pressure oil supply

Steam inlet

Governor intermediate pressure oil

Woodward PG governor

Trapped governor oil Pilot valve

Inlet steam valves

Drain oil to governor internal reservoir

Flyweights Pre-pilot valve

Accumulators

Oil pump Governor drain Governor drain

Lube system drain

High pressure oil from lube system

Mechanical–hydraulic governor system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)

© 2006 by Taylor & Francis Group, LLC

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FIGURE 8.7

Worm and wheel governor drive Inlet servo-motor

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Handbook of Lubrication and Tribology

the proper level and temperature. Water contamination, even in trace amounts, contributes significantly to early failure as well as forming oxides that also contribute to failures. As noted by one governor manufacturer, dirty oil causes most governor/actuator troubles [32]. Oil contamination and degradation is particularly problematic on self-contained sump units that do not have oil conditioners. Other parts of the governing system must also be maintained. Governor valves and linkages must be free from binding or sticking. Valve sticking may be related to steam impurity, which may lead to deposits on the valve stem [33], or it may be related to oil contamination causing deposits or corrosion in tight clearance hydraulic components [34], which may have clearance as small as 5 µm. Hydraulic actuators may operate on the valves through a linkage. Loose or worn linkage components can cause unacceptable governor control. Linkage bearings are usually hand-oiled or greased. Some, however, are made from low-friction materials, which may require little or no lubrication.

8.3.4 Turning Gear High-temperature steam turbines are sometimes equipped with a turning gear to prevent bowing of the rotor when at rest, especially after shutdown. The need for a turning gear depends upon the probability of rotor bow, which is related to the steam temperature, shaft diameter, and bearing span. Turning gears are primarily found on large turbines with long bearing spans, though they are sometimes needed for small turbines as well to allow for oil circulation through the bearings during cool down. The turning gear is operated prior to turbine run-up and immediately after shutdown. Turning gears are electric motor driven with a means for disengagement such as a clutch or retractable gear. The turning gear motor is typically grease lubricated whereas the actual turning gear and bearings are lubricated with oil supplied from the main circulation system. A separate, relatively small, motor-driven oil pump is generally provided to supply oil to the bearings of the turning gear system. The auxiliary oil pump, which backs up the main oil pump, may also be used for this service. During turning gear operation, oil inlet temperature may be kept cool to increase oil viscosity thereby maintaining a thick oil film in the turbine bearings during low-speed operation.

8.3.5 Couplings Couplings are used to connect the steam turbine to the driven equipment. They are made from corrosionresistant or coated materials. Couplings can be either rigid or flexible. Rigid couplings are essentially two flanges bolted together. Such couplings require no lubrication, but do not readily accommodate changes to machine position, which can be caused by thermal expansion of the equipment, foundation settling, and strain due to loading. Flexible couplings accommodate some misalignment; however, their use does not preclude the need for proper machine alignment of both the turbine and driven equipment [6]. Flexible couplings are described by a number of standards such as API 671, ISO 10441, and MIL-C-23233A. For turbine applications, special attention may be required with respect to machinery alignment due to thermal expansion. Quill shafts, membrane couplings, and contoured disc couplings run dry and without lubrication and are often preferred for their low maintenance. Gear couplings must be lubricated. 8.3.5.1 Gear Couplings Gear couplings can be advantageous because of their light weight and minimal required overhang, and because they allow for maximum axial movement between turbine and driven equipment shaft ends as caused by expansion of various parts under hot conditions [35]. In general, however, the need for lubrication and maintenance means that geared couplings are seldom used in new turbine applications though there is still a considerable population of geared couplings that must be maintained. The life of a geared coupling is primarily dependent on alignment and lubrication. The majority of geared tooth coupling failures are due to improper or insufficient lubrication [36]. Gear coupling lubrication is complicated by the centrifugal effect that a spinning coupling has on lubricants. Packed lubrication with grease can only be applied at relatively low speeds since the thickener tends to separate out of the grease under high

© 2006 by Taylor & Francis Group, LLC

Steam Turbines

8-15

centrifugal force [37]. Packed lubrication with either grease or oil may require lubricant replenishment at 6 to 12 month intervals [38]. Typical grease used for high-speed geared coupling is NLGI #1 or #2 grade with R&O inhibitors. Greases that are specifically formulated for high-speed coupling application use thickeners, which have a density closer to that of oil [39]. These formulations resist separation due to centrifugal effects. Special grease formulations can also extend replenishment intervals beyond the 6 to 12 months typically cited. A test method for evaluating grease separation is ASTM D4425. In this test, the grease is subjected to 36,000 G centrifugal acceleration at 50◦ C for a period of at least 6 h. Results from the test are presented as K 36 = V /H

(8.2)

where V is the oil separation in volume percent, and H is the accumulated time of testing in hours. High speeds and low maintenance in gear couplings require the use of continuous lubricant feed at each hub that is provided by either oil spray or jet using filtered oil piped from the system bearing oil supply. Coupling teeth for such applications are often hardened usually by nitriding [39]. Lubricants must be carefully selected with additives that resist separation from centrifugal force. Oil additives, in particular silicone antifoam compounds, can separate out of the lubricating oil and form sludge [40]. The coupling lubricant must also resist reaction with metal particles that may exist in the coupling due to wear [41]. Such measures are unnecessary when using dry couplings. The coupling housings provide safe enclosure of the coupling. Coupling housings may also act as part of the lube oil system. The housings are oil tight and include provision for coupling lube oil supply if needed and drainage back to the reservoir. The drains also handle any oil that may be carried over from the coupled equipment and are consequently featured on housings for both dry and lubricated couplings. A filter breather is attached to the coupling housing to allow proper drainage or the housing is connected to the bearing oil vent system of the equipment train. Regardless of the type of coupling used, proper design of coupling housings is important due to windage losses, heat generation, and potential for oil leakage from the joined equipment [42].

8.3.6 Additional Tribological Components and Issues Several components outside of the lubricating oil circuit require batch lubrication and special material consideration to limit wear and corrosion. Among these components is the turbine casing and steam patch components. Gland seals may also be subject to wear and have considerable effect upon water contamination of the lubrication system. 8.3.6.1 Casings Steam casings expand and contract due to changes in casing temperature caused by the use of high temperature steam. Thermal movement is typically accommodated at the steam end by either a flexible support or sliding pedestals. Sliding pedestals are most common on large turbines and rarely used on small and medium sized units. Sliding pedestals may operate dry, or they may be lubricated by either grease or oil depending on the load, temperature, and expansion. The use of lubrication reduces friction thereby allowing casing thermal expansion without binding. Binding of the casing can cause distortion, misalignment, and vibration. Grease lubricated casing supports are often used for large central station steam turbines. Grease may be supplied by either a common system or grease gun. The type and application should follow the manufacturer’s recommendation. NLGI number 1 or 2 grease of sodium, lithium, or sodium–calcium soap base have been used for lubricating sliding pedestals; in addition a mixture of graphite and cylinder or turbine oil mixed to a paste consistency has also been used [43]. Problems associated with grease separation have been noted on high temperature, heavy turbines used in central station applications due to the high temperature and heavy loads associated with these applications [44].

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Handbook of Lubrication and Tribology

8.3.6.2 Steam Path Parts that are located in the steam path consist of the steam chest, rotating blades, stationary nozzles, diaphragms’, seals, valves, and valve guides. These parts have tribological issues that must be solved without lubrication, which would contaminate the steam. Stationary nozzles and rotating blades are damaged by erosion and corrosion mechanisms. Turbine blades in particular suffer the most damage in turbines resulting in loss of power, efficiency, or operation. There exist many causes for wear through the steam path; however, three mechanisms are prevalent. These are (1) moisture impingement erosion, (2) erosion–corrosion, and (3) solid particle erosion. Moisture impingement erosion is caused by the presence of water droplets in the steam. When turbine blades operate in wet steam, the moist steam may cause blade erosion. Erosion is dependent on speed of the rotating blade, wetness of the steam, and blade design. Blade velocities can exceed 250 m/sec (825 ft/sec) at the tip. Moisture impingement erosion has been noted to be particularly problematic in the final stage of long multi-stage low-pressure turbines due to condensation. Allowable wetness is related to steam conditions, blade velocities, and design. In some cases, there is no need for a moisture limit. In other cases, 8% [45], 12% [46], or other moisture limit is used depending on the application. Moisture erosion also effects seals and can lead to degradation of performance and changes to the thrust loading [47]. Erosion–corrosion problems are caused by reactive steam chemistry. Steam of insufficient purity may cause deposits on the casing, nozzles, blades, seals, and sealing surfaces. These deposits may contain corrosive agents such as chlorine, which can attack the material used on these components. This results in eventual pitting and stress corrosion cracking [48]. Geothermal steam applications are known to have particularly corrosive steam with constituents of silica, sodium, ammonia, calcium, and sulfate. The acidity of geothermal waters can be very high with pH as low as 1.8 [49]. Solid particle erosion (SPE) is caused by entrainment of erosive materials in the steam. Solid particle erosion is traceable to exfoliated material coming from the boiler tubes, and in some cases, the steam leads. This type of erosion appears to be related to both the size of the unit and the pressure being employed. The solid particle erosive mechanism is most prevalent on large central station utility turbines; and it is rarely observed on small turbines operating under 540◦ C (1000◦ F) [50]. An important factor in each of these erosive wear mechanisms is the condition of the steam. For this reason, steam conditioning may be used for ensuring reliable operation. Monitors have been developed to quantify the particle loadings from the boiler [46]. Various separators and moisture removal devices may be employed upstream and inside of the turbine. Strainers are used to remove the largest particles and trap foreign objects. Some recommendations for steam purity are specified by NEMA for lowpressure turbines relating to the amount of dissolved solids, alkalinity, conductivity, and content of silicon oxide, iron, copper, sodium, and potassium [6]. Original equipment manufacturers also provide steam purity recommendations. Design methods for combating erosive wear include the use of either hardened materials or hard coatings such as Stellite on turbine blades. Stainless steels such as 12% chromium steel are also used. For turbine nozzles and internals, chromium steel cladding may be used [50]. Corrosion-resistant coatings have also been developed for this service. Other forms of wear and degradation internal to the steam path also exist. Turbines that are not in operation can experience a form of corrosion known as stand-by corrosion [45,49]. The corrosion is due to steam leaking into the turbine past a valve, which is not tight. Once the steam has leaked into the unit, it can condense and corrode the unit. This type of corrosion may cause severe pitting on stainless steel buckets. Brown specs (known as tubercles or scabs) form on carbon steel parts, such as discs and diaphragms [46]. It is therefore important that idle turbines have the inlet valve tightly seated and that all the casing drains be open [49]. Additional measures such as an additional drain between the turbine and steam inlet valve, and blanketing turbine internals with a positive flow of dry gas along with running the lubrication system and rotating the journals have also been performed [43]. 8.3.6.3 Seals and Gland System In order to maintain efficiency and performance, seals are required to limit steam leakage from the turbine case and between each stage. Casing end seals, also referred to as packing seals, are provided where the

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Steam Turbines

8-17

shaft ends pass through the casings. They are used to seal against the leakage of steam to atmosphere, and to seal against air suction into the low-pressure condensing case of condensing turbines. The seals may be carbon ring, labyrinth, or noncontacting mechanical design. Properly functioning gland seals are important for maintaining turbine performance. Improperly functioning gland seals can cause excessive steam leakage, which may ingress into the bearing housing leading to oil contamination. Likewise, a loss of vacuum on condensing turbines can also cause excessive steam leakage past the gland seals. Gland leakage does not have to be visible to cause a problem.

8.3.7 Driven Units Driven units such as gears, compressors, and generators create additional complexity to a lube oil system. Many of these units will have similar requirements to the turbine with respect to journal and thrust bearing lubrication. The oil selected for a common lube oil system must be suitable to all the pieces of equipment to be supplied. Low-viscosity rust and oxidation inhibited (R&O) oils, commonly called turbine oils, are used in many high-speed gear units where the gear tooth loads are relatively low [51] and the high entraining velocity of the gear develops thick elasto-hydrodynamic (EHD) oil films. Slower speed gears, as used for propulsion, tend to be more heavily loaded. These gears generally require higher viscosity lubricants with antiscuff additives [51]. High-pressure oil seals, as used in compressors; and hydrogen seals, as used in generators, can cause contamination of the seal oil by gas such that natural or vacuum degassing is required [31]. In some cases, a separate, isolated, lube oil system is used to provide seal oil due to potential contamination of the lubricating oil [52].

8.4 Lube Oil Systems Lube oil systems may be classified as either nonpressurized or pressurized systems. Nonpressurized lube systems consist of ring lubrication and are common on very small steam turbines. Larger turbines use pressurized lubrication.

8.4.1 Nonpressurized Oil Ring Lubrication Ring-lubricated hydrodynamic bearings are used where service is not critical or the steam turbine is spared. These bearings have the advantage of not requiring an external lube oil system thereby enabling steam turbine application where initial cost is a primary concern. Figure 8.8 shows an oil ring-lubricated journal bearing. The oil ring lubrication system employs metal rings to deliver oil to the turbine bearings. The rings are rotated by the journals carrying oil from a sump below the bearings to the top half bearing liners where it is fed into the clearance between the bearing liners and the shaft journals. Oil is drained from the ends of each bearing liner and returned to the bearing housing reservoirs to be cooled. Some of the supplied oil may be used to feed a rolling element bearing that is normally required in conjunction with ring-lubricated journal bearings for thrust positioning. The use of ring-lubricated bearings is limited with respect to load capacity, journal rotational speed, and by the need for cooling. Bearing housings may be double walled to allow water circulation to remove heat from the oil bath. Under conditions of high inlet steam temperature, the bearings can be damaged after shutdown because there is no longer oil circulation to carry heat away from the shaft, and a turning gear is sometimes used to continue the rotation of the shaft and subsequent oil ring lubrication. Ring-lubricated bearing housings are equipped with constant-level sight-feed oilers that maintain a constant reservoir oil level. A permanent indication of the proper oil level is clearly marked on the outside of the bearing housing. Low oil level in the housing will cause inadequate bearing lubrication. Excessively high oil levels can also be detrimental as it may restrict oil ring rotation also causing inadequate bearing lubrication. Housings for ring-lubricated bearings are provided with plugged ports positioned to allow visual inspection of the oil rings while the turbine is running.

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Handbook of Lubrication and Tribology Inspection plug Rotor shaft

Oil ring

Journal bearing

Oiler

Oil reservoir Cooling water Cooling chamber

Lubricating oil Cooling water

FIGURE 8.8 Ring-lubricated journal bearings. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.) TABLE 8.2

Lube System Design and Procurement Standards

Standard designation API 614 ASTM D4248 ASTM D6439

Standard title Lubrication, shaft-sealing, and control-oil systems and auxiliaries for petroleum, chemical, and gas industry services Design of steam turbine generator oil systems Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems

8.4.1.1 Ring Lubrication System Maintenance Proper oil level should be maintained at all times. Since oil ring lubrication systems have no means of filtering solids from the oil or removing water, periodic sampling and frequent oil changes are necessary to ensure a clean oil supply. The range of cooling water temperature must also be controlled to ensure good heat transfer without promoting condensation in the oil sump. To avoid condensation, the minimum inlet water temperature to the bearing housings should preferably be above the ambient air temperature.

8.4.2 Pressurized Lubrication Systems The pressurized lubrication system is essentially a closed loop system designed to provide an uninterrupted supply of cooled and filtered oil at the proper pressure to the bearings, control-oil system, shutdown system, and other components such as continuously lubricated couplings, as well as gears and seals on adjoining equipment. Oil consoles vary widely depending on the make, size, type, and purpose of the turbine and its adjoining equipment. Lubrication systems are designed according to the application, which may require the use of design standards such as those shown in Table 8.2. Proper lube system design is vital to machine reliability. A typical system is shown in Figure 8.9 and Figure 8.10. The oil is taken from the reservoir and passed through a cooler then filtered. The flow is split into two legs. One leg delivers high-pressure oil to a

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Steam Turbines

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FIGURE 8.9 API 614 lube oil console. (From Elliott Company, Jeannette, PA. With permission.)

common header for the governing and control mechanisms. A second leg delivers reduced pressure oil to a common header for the bearings. Bearing oil pressure is typically 100 to 125 kPa (15 to 18 psi), but may range from as low as 55 kPa (8 psi) in some systems to 345 kPa (50 psi) in others. Oil from the bearings and governor mechanisms will drain back to the reservoir. Figure 8.10 shows the following major components: • • • • • •

Oil reservoir Pumps and drivers Filters Coolers Control valves Piping

Additional accessories may include relief valves, transfer valves, accumulators, and instrumentation as shown in Figure 8.10. Not shown is the oil conditioning hardware. Each of the major components is described briefly along with instrumentation, commissioning, and system maintenance. 8.4.2.1 Oil Reservoir The reservoir is usually of rectangular shape, carbon steel construction with an interior coating of rust proofing paint. Solid stainless steel or stainless steel clad construction is also used. Normally the reservoir will have a sloping bottom to drain, clean out manways, gasketed openings, fill opening with strainer, oil level sight gage, and vent with weatherproof breather. The various oil levels as defined in the reservoir are shown in Figure 8.10. Depending on design requirements, the reservoir is sized to contain an amount of oil for anywhere from 3 to 5 min working capacity as measured from the minimum operating level. Large reservoir capacity enables disengagement of entrained air or gas and the settling of water and solid contaminants. A high and low oil-level indicator and alarm are usually provided. A free oil surface in the reservoir of at least 0.37 m2 /lps (0.25 ft2 /gpm) of oil is required to enhance air disengagement from the oil [53]. In addition, the oil reservoir is designed with a sloping bottom (1 unit in 24) such that supplementary water and dirt can accumulate at the area of the low point drain, and thus be drawn off during operation.

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Lube oil supply

8-20

High pressure control oil supply To oil reservoir To oil reservoir

Cooling water out

Oil return from units

A

A

PCV

Fill line

Cooling water in

PCV

TE

TI

Oil coolers VENT

Cooling water out

Vent

PCV

Main oil filters

TI PI

e

Purge

RV

RV

Minimum oil level and alarm level

PI

PI

TI

LS

LI Reservoir

Heater

Transfer valve 6-way

Relief/safety valve

Check valve

Pump suction strainer

Globe valve

Primary oil pump and driver

Change capacity (oil required for initial system fill) 800 U.S. gallons

Secondary oil pump and driver

Electric motor driver AC or DC power

Cooling fan Oil filter

Pump suction level

Inside bottom of reservoir

LIT

Tank drain TE

(3 min. retention)

Level indicator

PSH

Pressure switch, high

LIT

Level indicator transmit

PSL

Pressure switch, low

LS

Level switch

RV

Relief valve

Pressure control valve

TE

Temperature sensor

Pressure indicator

TI

Temperature indicator

LI

Orifice Gate valve Ball valve

FIGURE 8.10

Lube oil console P and I diagram.

© 2006 by Taylor & Francis Group, LLC

PCV Double pass, shell and tube heat exchanger

Concentric reducer PI

Handbook of Lubrication and Tribology

1/2” per foot minimum Oil to clairifier

Top of oil reservoir Maximum oil level Working capacity 360 gallons

Purge Fill conn w/strainer Oil from clairifier

PSL

Start secondary pump o setpoint (PSIG) falling press

Dirty oil drain

PDIT

Exhauster vent

Slop

PI

Drain

Cooling water in

Retention volume 600 gallons

Coolers and filter vents

Clean oil drain

B

B

Steam Turbines

8-21

Most pressurized lubrication systems are constructed with some provision for ventilation although some systems enjoy satisfactory operation with no such provision. Effective ventilation of the lubrication system enables the reduction of moist air that affects the service life of the turbine oil. The provision of adequate ventilation is also helpful in reducing foaming where trouble from this source is encountered [54]. The following methods of ventilation are commonly used: Natural ventilation, vacuum ventilation, or dehumidifier system [54]. Figure 8.10 shows a system equipped with a vapor extractor. The extractor pulls a slight negative pressure that should result in no more than −0.5 kPa (−0.07 psi) in the bearing housing to keep oil vapors from escaping, but without pulling in atmospheric contaminants. Reservoirs normally have a connection for an oil conditioning system. Such oil conditioners can provide further purification by removing water, acids, and other contaminants not removed by the filters. These are discussed in more detail under oil maintenance. 8.4.2.2 Pumps and Drivers Two or more oil pumps are normally supplied with the lube oil system. One pump is considered the main oil pump and the other, the auxiliary. The pumps are sized with additional flow capacity to provide a positive flow of oil under all normal operating conditions and most abnormal conditions to the turbine and the driven equipment. Additional smaller pumps may be used to supply oil for special purposes such as turning gear operation; hydrostatic lift oil for highly loaded bearings; seal oil for hydrogen-cooled generators; or oil transfer through filters [28]. Positive displacement pumps have relief valves located at each pump discharge line to protect the pumps and system against excessive pressure. The main oil pump can be driven off the main turbine shaft, by an electric motor, or by a small steam turbine. Most often, the auxiliary oil pump is driven from a different source of power than the main oil pump. The auxiliary pump driver is selected to reflect availability of power or steam under emergency conditions. Should the main pump fail, the auxiliary pump will automatically start. If the pressure continues to drop, the turbine and driven equipment will shut down. Emergency situations where both pumps fail are handled by either an emergency oil pump sized to provide last-resort lubrication for coastdown or a rundown tank that provides lube oil by gravity flow. Rundown tanks are common in marine applications. On lube systems where the auxiliary oil pump is driven by a small steam turbine, an accumulator is incorporated into the system. The accumulator will maintain the required oil flow while the turbine (auxiliary pump driver) is accelerating to speed preventing a system shutdown in case of main pump failure. 8.4.2.3 Filters Twin filters with multiple cartridge filtering elements are normally used in the lubricating oil system. Filters are operated in the full-flow mode such that all oil being circulated to the turbine passes through the filter. Using two filters permits filtering element changes while the equipment is in operation. Filtration ratings should be a minimum of 25 µm, and filtration of 10 µm is typically required. Filters are sized for a maximum pressure drop of 35 kPa (5 psid) when clean and passing oil at the design temperature. Filtering elements are typically replaced when pressure drop reaches approximately 100 kPa (15 psid) above original clean value [55]. The effect of water on the filter must be considered. Water and corrosion-resistant filter cartridge materials are preferred. Such water-resistant filter cartridges should not deteriorate even if water contamination reaches 5% by volume and an operating temperature as high as 70◦ C (160◦ F). Depth type elements (e.g., cotton and nylon) can suffer from a phenomenon termed “cartridge erosion,” where oil velocity enlarges or erodes the filter passages over time, which effectively invalidates the filtration rating of the element [55]. Cartridge erosion problems are eliminated by conservatively replacing filter elements every 6 months. Filters are also replaced if the pressure drop from clean increases by 100 kPa (15 psid). Recommendations of the filter element manufacturer should be considered.

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8.4.2.4 Coolers Oil coolers are usually conventional shell and tube heat exchangers with removable tube bundles. Normally, water flows through the tube bundle thus allowing for easy waterside cleaning. Oil flows through the shell side in a single pass. Coolers are usually operated so that the oil is at a higher pressure than the water, thus, reducing the severity of water contamination caused by tube failure. Typical cooling requirement is to cool oil to 120◦ F. Where conditions do not lend themselves to water-cooled heat exchangers, such as desert or subzero installations, air blast oil coolers must be considered. Coolers may be used for heating during initial oil system installation and cleansing and it is important that the system be designed for such use if desired [56]. Problems with maintaining oil temperature could be caused by improper venting, malfunctioning temperature regulators, incorrect water pressure, or badly fouled coolers [28]. Tube failure may be caused by fatigue and erosion. Excessive water flow can cause flow-induced vibration of the cooling tubes, but maintaining proper flow will reduce fatigue related problems. Water treatments and sacrificial anodes are used to retard corrosion failure of the cooler [56]. Cooler failures are responsible for the worst water contamination of the turbine oil. 8.4.2.5 Control Valves The backpressure regulator is designed to maintain a constant header pressure for all operating conditions. Normally, it is a self-operated valve, but where wide control ranges are required, pneumatic regulators complete with valve positioners are used. In addition to the backpressure regulator, pressure-reducing valves are required for all pressure levels below main header pressure. These valves are normally self-operated reducing valves but where wide control range is required, pneumatic operators complete with valve positioners may be furnished. 8.4.2.6 Piping Lubrication system components are joined together by the necessary piping to make the system functional. This includes provisions for the mounting of control instrumentation, such as pressure gauges, temperature gauges, switches, and monitoring and safety devices. Piping may be either carbon or stainless steel. Stainless steel is preferred due to superior corrosion resistance and is used extensively in refinery applications. The header piping connects the lube oil console to the various components being lubricated, such as bearings and seals. Used oil is returned to the reservoir through drain piping. Oil drains are sized to run no more than half full when flowing at a velocity of 0.3 m/sec (1 ft/sec) and are arranged to ensure good drainage. Horizontal runs slope continuously, at least 40 mm/m (1/2 in./ft), toward the reservoir [57]. 8.4.2.7 Safety and Monitoring Devices Lubrication system instrumentation is located throughout the system as shown on the schematic oil flow diagram. Monitoring devices, such as pressure gauges, safety devices, and alarm and trip switches are generally mounted on header piping close to the components being lubricated. A low-pressure start-up switch signals the auxiliary pump to start if pressure is too low. Temperature indicators are provided at bearing and seal outlets and at the inlet and outlet of coolers. Pressure indicators are generally provided at each pressure level. A sight-flow indicator is provided at the outlet of each turbine shaft bearing and each turbine thrust bearing. 8.4.2.8 Cleaning and Flushing All reasonable effort must be made to limit the introduction of contaminants into the lube oil system during construction. Proper cleaning and preservation of lube system components must be performed prior to system shipment. Different preservatives are used depending on the environment and expected storage time [58]. All units employing forced-feed oiling systems should have the entire lubrication system thoroughly flushed before operation. The importance of this step cannot be overemphasized. All dirt, rust scale, weld

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slag, or other contaminants that have been introduced into the oil system during storage, transportation, and fabrication at the jobsite must be removed by a continuous flushing operation or in extreme cases that do not involve stainless steel pipe, by pickling and cleaning. In addition to flushing during the commissioning, the system should also be flushed if left idle for a long time. Most turbine manufacturers provide special instructions for the oil flush. In the absence of such instructions, industry recommendations should be consulted such as those detailed in ASTM D6439 [59] or API RP 686 [58]. Flushing the system may require the use of external pump and such preparations should be made in advance. The bearings and bearing area should be bypassed until the system is proven to be clean. The flushing should continue until the required cleanliness is achieved based on inspection of the flushing filters or strainers, patch test, particle counters, or ISO 4406 cleanliness level. Flush oils, operating oils, and preservative oils must be compatible to preclude foaming, formation of emulsions, or breakdown of oil additives. Compatibilities and limitations may generally be obtained from the oil supplier. A system that is to use phosphate ester fluids must be flushed with phosphate ester fluid since such fluid is incompatible with mineral oil. The same may apply to other synthetic oils. 8.4.2.9 Lube System Maintenance Lube systems must be periodically inspected and maintained to ensure their proper operation. As a minimum, the following regular checks should be performed: • Check filter pressure drop and replace elements as recommended. • Check the oil reservoir level and add oil as required. • Periodically check operation of auxiliary oil pump by operating pump and returning to auxiliary duty. In addition, turnaround maintenance of the lube oil system should be performed at 1 to 3 yr intervals, as normal plant maintenance permits. Care must be taken to keep contaminants out of the lube oil circuit during bearing changes, filter changes, top up, and other maintenance activities.

8.5 Turbine Oil Equipment vendors often have turbine oil standards detailing the minimum characteristics required for successful turbine operation. In the absence of such standards, an internationally recognized turbines oil specification such as shown in Table 8.3 should be used.

8.5.1 Physical Properties Turbine oil performs four functions (1) Lubricate bearings and gears; (2) cool lubricated parts, carrying heat away from hot surfaces such as bearings and shafts; (3) act as a hydraulic fluid for governor, control valves, and safety devices; and (4) act as a sealant for gas seals such as hydrogen shaft seals in generators or gas seals on compressors. Each of these functions require an oil that is suitable with respect to several physical, chemical, and performance properties. Some physical properties frequently used to characterize turbine oils with corresponding American Society for Testing and Materials (ASTM) test methods TABLE 8.3

Standards for Turbine Oils and Hydraulic Fluids

Standard designation ASTM D4293 ASTM D4304 ISO 8068 MIL-PRF-17672D MIL-PRF-17331H

Standard title Standard specification for phosphate ester-based fluids for turbine lubrication Standard specification for mineral lubricating oil used in steam or gas turbines Petroleum products and lubricants — petroleum lubricating oils for turbines (categories ISO-L-TSA and ISO-L-TGA) Performance specification: hydraulic fluid, petroleum, inhibited Performance specification: lubricating oil, steam turbine and gear, moderate service

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8-24 TABLE 8.4

Handbook of Lubrication and Tribology Standardized Lubricating Oil Analytical Techniques

Physical or chemical property

ASTM designation

ISO designation

ISO viscosity grade Kinematic viscosity at 40◦ C, 100◦ C Viscosity index

D2422 D445 D2270

ISO 3448 ISO 3104 ISO 2909

Pour point Flash point Total acid number (TAN)

D97 D92 D974

ISO 3016 ISO 2592 ISO 6618

Foaming characteristics Air release

D892 D3427

ISO 6247 DIN 51 381

Water separability (Demulsibility) Rust prevention

D1401 D665

ISO 6614 ISO 7120

Corrosiveness to copper

D130

ISO 2160

Oxidation stability (TOST) Rotating pressure vessel oxidation test (RPVOT) Acid number Karl Fischer titration Color ISO cleanliness code

D943 D2272

ISO 4263

D664 D1744 D1500

ISO 6619 ISO 6296 ISO 2049 ISO 4406

Purpose Overall viscosity classification Relates to viscosity at normal operating conditions Empirical comparison of viscosity and temperature characteristics Measures low temperature flow properties Low value indicates volatile components Determination of acidity of new and used oils by titration with KOH Foaming characteristics of lubricating oils The oil’s capacity to separate entrained air over a period of time Emulsion characteristics of oil Ability of oil to prevent rusting of steel surfaces in presence of water Indicates tendency of oil to corrode copper and copper alloys Oxidation stability of mineral oils Tests remaining oxidation life of in-service oils Indicates acid level Measures the water content of oil Measures color Measures oil cleanliness

are summarized in Table 8.4. Detailed descriptions of the ASTM methods are available in the ASTM Handbook [59]. The most important physical property is viscosity. Table 8.5 gives the viscosity ranges for typical mineral lubricating oils used in steam turbines. Typical viscosity grade numbers are ISO-VG-32, VG-46, VG-68, VG-78, and VG-100 such that the viscosity grade numbers indicate the average oil viscosity in centiStoke units at 40◦ C (104◦ F). In order to reduce the power losses at the bearings and improve the responsiveness of hydraulic components, the lowest acceptable lubricant viscosity is normally selected. As a result, the usual lubricant employed in a common oil system is ISO VG-32 turbine oil cooled to a supply temperature of 120◦ F after the cooler. Other viscosity grades are also used. ISO VG-46 turbine oil cooled to a supply temperature of 140◦ F after the cooler is commonly used in desert, arid, and offshore applications where air blast coolers are utilized, or where the ambient temperature is quite high [56]. Oils used for ringoiled turbine bearings tend to be higher viscosity such as ISO VG-68 or VG-100. Oils used for shipboard propulsion may be ISO VG-68 to VG-100 and may have mild antiscuff additives. It is important to note that the lube oil system and the turbine rotordynamics are designed considering a specific oil viscosity. Turbine lube systems must be maintained with lubricants of the recommended viscosity, and the viscosity specification should not be changed without proper engineering review.

8.5.2 Formulation To achieve the desired physical, chemical, and performance properties, turbine oil is formulated with a base fluid and additive package consisting of rust and oxidation (R&O) inhibitors. Steam turbine oils are essentially special grades of R&O oils, formulated to give better oxidation resistance and longer life in a steam turbine [60]. While industry standard lube oil bench tests can provide great insight into the performance and life expectancy of turbine oils, both turbine original equipment manufacturers (OEMs) and oil suppliers generally agree that past successful performance of a particular oil under similar conditions is the best overall representation of quality and performance [61].

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Steam Turbines TABLE 8.5

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Physical Requirements for Turbine Oils

ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. Military specification Military symbol ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. At 100◦ C Pour point, ◦ C, max. Flash point, ◦ C, min. Viscosity index, min. Total acid number (TAN), mg KOH/g, max. Corrosiveness to copper, max. Rust prevention Water, percent Valve sticking characteristics Foaming characteristics Sequence 1, mL max. Sequence 2, mL max. Sequence 3, mL max. Air release Water separability Oxidation stability, min.

Light turbine oil

Medium turbine oil

Medium-heavy turbine oil

Heavy turbine oil

Test method

32

46

68

100

D2422 D445

61.2 74.8

90 110 MIL-L-17331J 2190 TEP

28.8 35.2 2075 T-H 32

41.4 50.6 MIL-L-17672D 2110 T-H 46

2135 T-H 68

D2422 D445

28.8 35.2 Report −29 157 94 0.20

41.4 50.6 Report −23 163 94 0.20

61.2 74.8 Report −18 171 94 0.20

74 97 8.0 −6 204

1 Shall pass None Shall pass

1 Shall pass None Shall pass

1 Shall pass None Shall pass

1 None None Shall pass

65/0 65/0 65/0

65/0 65/0 65/0

65/0 65/0 65/0

40/40/3 1000 h

40/40/3 1000 h

40/40/3 1000 h

65/0 65/0 65/0 20 40/—/3 1000 h

0.3

D97 D92 D2270 D974 D130 D665 D95 D892

D3427 D1401 D943

8.5.2.1 Base Oil The base oil stock of a turbine oil comprises more than 98% of the formulation. The base oil is categorized as either conventional solvent refined mineral-based (API Group I), or hydroprocessed mineral-based (API Group II) oil. Group II base oils contain fewer heteroatoms (sulfur, nitrogen, oxygen), and have less aromatic content than Group I base oils. When properly formulated, Group II turbine oils will have longer oxidation life, less deposit forming tendencies, improved water shedding ability, and overall higher performance than do Group I turbine oils [60]. One advantage of the conventional mineral-based (Group I) turbine oils is better innate solvency than the hydroprocessed (Group II) oils. The better solvency of the Group I turbine oils provides better additive package retention and increased ability to dissolve oxidation products that could otherwise potentially lead to varnish and sludge. While Group I and Group II base stocks are compatible with each other, the additive packages used to formulate the respective turbine oils may be incompatible with the overall mixture. Mixing oils can therefore cause sludge formation and additive dropout [62]. For this reason, compatibility between products is an important consideration when mixing two oils. 8.5.2.2 Additives Additives are used to improve the performance of the oil. Although additives are to some extent consumed in performing their functions, they can be replenished through normal lubricant make-up thereby enabling suitable performance for longer periods. Note that newer machine designs offer less oil loss and therefore do not benefit as much from this effect as did older machines exhibiting greater oil loss. The main types of additives include oxidation inhibitors, rust inhibitors, foam inhibitors, and demulsifiers.

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8.5.2.2.1 Oxidation Inhibitors Antioxidants are the additives, which have the strongest influence on the useful life of turbine oils. They generally function either by free radical inhibition, by hydroperoxide decomposition, or by deactivation of metal catalysts. The two major types of antioxidants used in turbine oils are arylamines and hindered phenols [63], and they work as free radical inhibitors. A mixed phenol-amine has certain advantages over the use of a single antioxidant system. Other additives and combinations of additives are also used to suppress oxidation. In particular, metal deactivators are used to suppress oxidation by reacting with metal ions and surfaces to inhibit their catalytic activity [64]. 8.5.2.2.2 Corrosion Inhibitors Highly refined oils lose their metal-wetting ability and are easily displaced by water. For this reason, corrosion inhibitors are necessary to prevent corrosion. New turbine oils contain a rust-inhibitor additive and must meet ASTM Test Method D 665. These corrosion inhibitors typically work based on the physical adsorption principle. In action, the corrosion inhibitor “plates out” on surfaces, forming a film that resists displacement by water and, therefore, protects the surfaces from contact with water [65]. Corrosion inhibitors used in turbine oils are polar and thus susceptible to water washout. Alkenyl succinic acids are therefore widely used due to their resistance to water washout [66]. 8.5.2.2.3 Foam Inhibitors Foam additives must be carefully selected in order to prevent excessive foam formation, but still retain short air release times [67]. Highly refined hydrotreated base oils have lower foaming tendencies than conventionally refined base oils. Foam inhibitors work by decreasing the gas-lubricant interfacial tension. Liquid silicones are an effective antifoamant, but also act as an air-emulsion stabilizer, negatively influencing the air release properties of the turbine oil as it resides in the stilling portion of the equipment. 8.5.2.2.4 Demulsifiers Demulsifiers destabilize oil–water emulsions by changing the interfacial tension of oil and water thereby allowing their separation [64]. Conventional mineral-based (Group I) turbine oils usually contain demulsifying additives whereas hydroprocessed (Group II) turbine oils have good demulsibility without an additional additive.

8.6 Performance Features of Turbine Oils The following oil performance features must be retained to ensure safe and continuous operation of the turbine (1) viscosity; (2) oxidation stability; (3) freedom from sludge; (4) anticorrosion protection; (5) water separability [68]; and (6) air separability and resistance to foaming.

8.6.1 Viscosity Viscosity is measured by ASTM Test Method D 445. Viscosity is the most important characteristic of turbine oil, as the oil film thickness under hydrodynamic lubrication conditions is critically dependent on the oil’s viscosity characteristics. Viscosity also affects journal bearing stiffness and damping properties, which determine the vibration characteristics of the turbine. Viscosity of most new oils may vary by ±10%. A change in viscosity up to 10% is not in itself likely to cause trouble; however, a change in viscosity of 5% from its original value should be investigated for the cause. A change in viscosity is usually caused by contamination or top off with the wrong lubricant rather than by degradation of the oil. Drop in oil viscosity is a particular concern where turbine driven compressors are used in the compression of hydrocarbon gases because the viscosity change may be caused by contamination of the oil from the lighter hydrocarbons [56].

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8.6.2 Oxidation Stability During steam turbine operation, the lubricating oil is subjected to relatively severe oxidizing conditions. These are due primarily to the influence of heat, the presence of water and entrained air, and the catalytic action of substances in contact with the oil, particularly copper and ferrous metals [69]. Under these influences, the antioxidants are gradually used up and the oxidation stability decreases. ASTM D943 Turbine Oil Stability Test (TOST) is used to evaluate the oxidation characteristics of new inhibited steam turbine oils. This is an accelerated oxidation test; actual service should be much longer than test report hours [61]. Since TOST testing can take longer than a year, it is impractical as an in-service oil test.

8.6.3 Freedom from Sludge and Deposits Deposits are generally formed due to oxidation of the turbine oil, soap formation, microbiological growth, contamination by water containing salts, and solid particulate contamination [66]. Process gases can also react with the oil and its additives to form deposits. One such example is a turbine driven ammonia compressor in which the oil became contaminated with ammonia. The acidic rust inhibitor used in the turbine oil reacted with ammonia to form an insoluble resinous product [70]. Filtration and centrifugation can remove sludge and other products from oil as they are formed, but if oil deterioration is allowed to proceed too far, sludge will deposit in parts of the equipment and system flushing and an oil change may be required [71].

8.6.4 Corrosion Protection Protection against rusting is very important due to the common presence of water in turbine oils and water vapor in the ambient air. Rusting may occur below the oil surface, at the oil surface, or in the vapor spaces above the oil surface. Rusting requires oxygen, water, a corrodible surface, and time. Effective corrosion protection requires the elimination of any one of these items. Oil acts to protect against corrosion by coating structural surfaces with corrosion inhibitor thereby denying access of water to corrodible surfaces. ASTM D665 is used to evaluate the rust-preventing characteristics of steam-turbine oil in the presence of water. Procedure A is used for land turbines where condensed steam or humidity from air is the water source. Procedure B is used for marine-service ocean-going vessels where salt water can be a water source. Present additive technology has been found to be highly effective at preventing rusting problems below the oil surface in full flow conditions. When rusting occurs below the oil surface, it is frequently caused by galvanic corrosion, and it is noticed in areas where there is little oil movement and where free water collects, such as the bottom of the oil reservoir. Galvanic corrosion is caused by contaminant particles settling out of the oil and the presence of water. Particulate matter can create galvanic cells and act as nuclei for air bubbles [34]. Factors that influence galvanic corrosion are impurity concentrations, the pH of the water, and temperature [66]. Galvanic corrosion shows up as black rust. Rusting at the oil surface is typically caused by liquid water standing on the surface [72]. Most rust problems occur above the oil in what is known as the vapor space. Vapor spaces are present in steam turbine bearing pedestals, oil return lines, sumps, and gear cases. The air in these vapor spaces will contain water vapor from the relative humidity of the air drawn into the system and from the evaporation of water entrained in the oil. In addition, salt particles that can act as corrosion-sponsoring nuclei also may be present [73]. Water vapor tends to condense on the cooler parts of the circulation system, such as the underside of the reservoir top, inside return-oil piping above oil level, in bearing pedestals, and around governor parts [74]. Corrosion in the vapor space results in formation of scaly red rust.

8.6.5 Water Separability (Demulsibility) A lubricant’s ability to separate readily from water is one of the most important requirements of a turbine oil. Water must readily separate from oil in the drain tank so that it is dry when pumped to the system.

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Demulsibility is influenced by oxidation and contamination from dirt or metallic particles. Resistance to oxidation helps preserve the demulsibility characteristics of the oil. Normally, if the oil is in good condition, water will settle to the bottom of the storage tank, where it should be drained off as a routine operating procedure [71]. Water may also be removed by purification systems. If turbine oil develops poor demulsibility, significant amounts of water will stay in the system and create problems such as increased oxidation, additive depletion, and corrosion. ASTM D1401 is used to test the demulsibility characteristics of oil.

8.6.6 Air Separability and Resistance to Foaming All oils will foam in some degree. Foaming of the present day turbine oils should not, however, occur unless the oil is contaminated or subjected to abnormal aeration. Antifoam additives suppress foam, but in doing so may also slow down air release leading to air entrainment. Air entrainment in the oil has been known to cause pressure surge in oil systems, interruptions in oil supply, excessive formation of foam [75], and reduced hydraulic control. Care must be taken such that improving the antifoam characteristics of turbine oil does not lead to unacceptable air separability characteristics. Turbine circulation systems have been constructed to eliminate conditions that have been found to cause foaming such as leaky pump suctions, excessive splashing of oil returning to the reservoir, oil-return lines of insufficient size or capacity, and insufficient venting. Wide differences in temperature between the fresh oil (as added) and the oil in the system may contribute to foaming [76]. Serious cases of excessive foaming may be due either to mechanical faults of the type listed or to oil contamination [77]. Problems with excessive foaming may also be due to mixing of incompatible lubricants [63] or the use of excessive antifoam inhibitor. Air entrainment issues are also affected by system design. In particular, the stilling period of the lube oil system can affect the air entrainment characteristics of oil. Machines that provide short stilling periods for the oil have displayed air entrainment/release characteristics that seem to counter those displayed during standard air release testing (ASTM D3427). Such machines with very short stilling periods have displayed increased air entrainment when nonsilicone antifoamants have been used and it is suspected that the silicone antifoams discourage the initial air entrainment during the agitation period [78].

8.7 Degradation of Turbine Oils in Service Factors responsible for oil degradation in service include contamination, additive depletion, oxidation, and bacteriological deterioration.

8.7.1 Contamination Contaminants will unavoidably find their way into the lubricating oil. The following types are most common: water, oil soluble contaminants, and solid particles. 8.7.1.1 Water Water is always present in oil in solution and may also be present in free or emulsified form. The solubility of water in oil is temperature dependent. Water in solution has no adverse effect on lubricating properties and will not cause corrosion; however, when hot oil subsequently cools, some water may come out of solution as very fine droplets dispersed throughout the oil [28]. This water is very likely to cause corrosion of steel parts and may also cause other problems, (e.g., foaming, sludge formation, and change of viscosity). In addition, water can also lead to oxidation, additive removal, bacteriological contamination, as well as reducing filter element life. Water enters the oil system from the condensation of humid air by system temperature fluctuation; from steam through the turbine gland seals; or from leaking oil coolers [73]. Leaking gland seals is the

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most common source of water. Leaking oil coolers is the most detrimental particularly since cooling water leaks will have moderate to high concentrations of dissolved solids. In extreme cases, a rupture of the heat exchanger can cause massive amounts of water to enter the machine compartment [79]. Free water generally exists above a saturation level of around 120 to 150 ppm, and oil becomes cloudy in the range of 200 to 500 ppm [80]. A centrifuge is effective in removing free water down to about 30 ppm above the saturation level. Different methods for the testing of water exist. The simplest is visual inspection followed by “crackle” or hot-plate test, which can indicate the presence of water in oil due to boil off. Another test is the Fourier Transform Infra-Red (FTIR) spectrometry. The Karl Fischer Titration, ASTM Test Method D 1744, is the most accurate method for water testing. Differing limitations for water are noted by different manufacturers. In general, the water content should never be allowed to exceed 2500 ppm (0.25%). ASTM D4378 cites 1000 ppm (0.1%) as a warning limit. Depending on the design and application, some manufacturers will require a limit of 500 ppm (0.05%). 8.7.1.2 Soluble Contaminants Oil soluble contaminants may include gases, solvents, other lubricants, flushing oils, preservatives, and sealants. Gases and some light solvents can be removed by vacuum dehydration methods. Other contaminants cannot be removed. The presence of such contaminants requires the consultation of the oil supplier and the turbine manufacturer. A common source of dissolved gases is the oil seals used in some generators and compressors. 8.7.1.3 Solid Particles Abrasive contaminants can damage bearings, journals, and control mechanisms. Improved practices such as better preservation of the turbine and its components when not in operation, high velocity system flushing during commissioning, and use of full flow filtration during operation have led to a significant reduction in failures due to abrasive contaminants [81]. Cleanliness of the system oil can be determined by gravimetric means by ASTM F 311 or F 312 or by particle counting. Allowable contamination level is dependent upon the individual turbine application and components in the system. ISO 4406 cleanliness levels ranging from 18/16/13 to 16/14/11 are commonly applied to steam turbine service. The three digits of the ISO 4406 code refer to the number of particles per milliliter greater than 4-, 6-, and 14-µm respectively. It should be noted that further reductions in contamination beyond manufacturer recommendations might lead to improvements in reliability that can be cost justified [82].

8.7.2 Additive Depletion Additives are used up in the performance of their function. In other cases, the additives are removed due to reaction with contaminants or drop out due to problems of compatibility. Oil suppliers are often able to replenish additives by sweetening the oil.

8.7.3 Thermal and Oxidative Degradation The oil acts as a heat transfer fluid with the overall system design determining the heat load on the oil. Factors such as smaller oil reservoirs, higher shaft surface velocity, and higher shaft and bearing temperatures all contribute to environmental conditions that degrade the oil by thermal stress leading to oxidation. Oxidation occurs by chemical reaction of the oil with oxygen. The first step in the oxidation reaction is the formation of hydroperoxides. Subsequently, a chain reaction is started and other compounds such as acid, resins, varnishes, sludge, and carbonaceous deposits are formed [71]. Oxidation products may further lead to rust and corrosion, and promotion of foaming and poor demulsibility. The oxidation rate is influenced by the presence of water, contaminants, entrapped air, and temperature. The oxidation rate of a fully inhibited mineral oil is quite low at temperatures less than 60◦ C and will double for every 10◦ C rise in temperature [83].

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For in-service oil testing, the oxidation stability reserve is best determined by the rotating pressure vessel oxidation test (RPVOT), ASTM test Method D 2272, and by total acid number (TAN), ASTM Test Method D 974.

8.7.4 Biological Deterioration Lubricating oils are susceptible to biological deterioration if the proper growing conditions are present. Procedures for preventing and coping with biological contamination include cleaning and sterilizing, addition of biocides, frequent draining of moisture from the system, and avoidance of dead-legs in pipes [71]. Sustained high water content can lead to bacterial and fungal growth in the system. This can cause filter blocking and formation of deposits. The most effective antimicrobial measure is the establishment of preventative procedures such as frequent draining of free water from the oil reservoir. Biocides are used to prevent microorganism growth. Sterilization by heat is also effective.

8.7.5 Turbine Oil Severity The expected service life of a turbine lubricant depends considerably on the severity of the application. Many low severity steam turbines have a history of requiring a full lubricant changeout only every 10 to 20 yr or longer, with periodic top-up with fresh oil [67]. Certain environmental conditions, however, can result in or accelerate lubricant degradation and reduce life. As noted, factors responsible for oil degradation in service include contamination, additive depletion, and thermal, oxidative, or physical breakdown. Other important factors affecting service life are (1) type and design of lubrication system, (2) condition of the system after construction, and (3) oil makeup rate. These factors vary from unit to unit so that service life is difficult to predict solely on original oil properties [84]. One method for determining the service conditions for each operating unit is to use a property called the turbine severity level (B), which is defined as the percent of fresh oil oxidation resistance or oxidation inhibitor lost per year due to oil reactions [85]. The equation for turbine severity is B = M · (1 − X /100)/(1 − e−M ·t /100 )

(8.3)

where B is turbine severity, % of fresh oil oxidation resistance lost per year due to oil reactions in the turbine, M is fresh oil makeup, % per year, t is years of oil use, and X is used oil oxidation resistance by ASTM D2272, % of fresh oil. A lubrication system with a high severity level requires frequent makeup or completely new charges, whereas one with a low severity level may have no problems with routine makeup [86]. The method requires periodic testing of the lubricating oil. Large steam turbines should have their turbine severity determined. The severity constant is different for each turbine, and varies widely between 5 and 30 for large turbines [87]. Figure 8.11 shows the importance of makeup rate for maintaining oil quality in a high-severity turbine where B = 25%/yr [85].

8.8 Lubricant Maintenance Small turbines with ring-lubricated bearings, and governors with sumps require periodic changes in lubricant. The quantities of oil are small, and it is often more economical to change the oil rather than to maintain it. Change periods of 1 yr are not uncommon and are set by regular change intervals, by monitoring the acid number or by more sophisticated monitoring. In larger turbine systems that employ circulating oil systems using more than 200 l (roughly 50 gal) of oil that require long periods of continuous operation, oil analysis generally proves more profitable than a routine time/dump program [88]. A life of up to 30 yr is desirable because of the outage and oil change

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100 90 80 70 60

40 30 M =3

0%

20 M 5%

=2

0%

10 9 8 7

M=2

M = 10%

M = 0%

Percent of original oil life remaining, X

50

6 5 0

4

8

12

16

20

Years of use

FIGURE 8.11 Effect of makeup rate on oil degradation for turbine severity, B = 25%/yr. (From DenHerder, M.J., Vienna, P.C., Lubrication Engineering, 37, 67, 1981. With permission.)

TABLE 8.6

Standards for Turbine Oil Maintenance

Standard designation ASTM D4378 ASTM D4057 ASTM D6224 ASTM D6439 IEC 60962 IEC 60978

Standard title Standard practice for in-service monitoring of mineral turbine oils for steam and gas turbines Standard practice for manual sampling of petroleum and petroleum products Standard practice for in-service monitoring of lubricating oil for auxiliary power plant equipment Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems Maintenance and use guide for petroleum lubricating oils for steam turbines Maintenance and use guide for petroleum lubricating oils for triaryl phosphate ester turbine control fluids

costs involved. In such systems, regular sampling and testing can indicate the need for oil conditioning. Many oil suppliers offer programs to meet specific lubrication maintenance requirements. Standards for turbine oil maintenance are listed in Table 8.6. Such standards offer a guideline for oil-monitoring and maintenance. Other methods may be applied depending on the application. One such standard, ASTM D4378, Standard Practice for In-Service Monitoring of Mineral Turbine Oils for Steam and Gas Turbines is used in the power generation industry [89]. As with any oil monitoring program, proper sampling is important. In-service oil should be tested at sufficient intervals to detect contamination, oxidation, and additive depletion. Key tests include appearance and color, water content, viscosity, total acid number, rust test, cleanliness, and RPVOT [89]. Systems that are exposed to volatile gases or liquids

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may also benefit from flash-point testing. Maintaining the lubricant may require new oil makeup, lube oil conditioning, and refortification.

8.8.1 New Oil Makeup Oil is lost due to leakage and to system maintenance such as draining off impurities, and filter changes. There is considerable variation with respect to the amount of makeup oil required for a steam turbine. Makeup rates can range from less than 5%/yr to more than 30%/yr in extreme cases. The average makeup rate in the United States is 7 to 10%/yr [89]. The compatibility of the system oil and the makeup oil are of critical importance. Compatibility is described as a lubricant’s ability to be mixed with another lubricant without detriment to the properties and the characteristics of either lubricant. The introduction of Group II oils has caused some concerns with respect to compatibility with Group I oils. In particular, the different additives and the solubility of those additives is a concern when mixing different oils especially those involving different base stocks. Problems involving excessive air entrainment, varnish particle build up, development of sludge, sticking of governor proportional valves, and plugging of governor filters has been noted on hydroturbines operating on turbine lubricants [90]. In some cases, a complete system flush may be required to introduce a new oil. The use of makeup oil that is the same oil as is already in the system is preferred for the elimination of compatibility issues.

8.8.2 Lube Oil Purification All circulating lube oil systems use filters to remove particle contaminants and purify the oil. Devices for removing liquid contaminants such as water will also improve system reliability. The most common devices for removing liquid contamination are settling tank, centrifuge, coalescing filter, and vacuum dehydrator. The settling tank works best on a batch basis. The centrifuge, coalescing filter, and vacuum dehydrator are applied continuously with 10 to 20% of the volume of oil in the turbine system every hour. Systems of this type tend to remove impurities as fast as they enter the oil, thereby avoiding accumulations. Settling tank — Oil contaminants that are heavier than oil can be separated by gravity alone. Such settling is best accomplished in a settling tank that is separate from the main oil tank. Settling times can be very long and the results are often less adequate than the onstream methods. Centrifuge — In a centrifugal purifier, or centrifuge, centrifugal force is used to accomplish the separation of contaminants heavier than oil. A separating force several thousand times that of gravity is produced by rotating the liquid at 7,000 to 15,000 rpm. The centrifuge is particularly effective in removing water and larger, heavier particles of solid impurities. The extent to which extremely fine solid particles are removed depends on the rate of throughput and other factors. Centrifuges are capable of removing free water and solids. Coalescing filter — A coalescing filter system uses special cartridges to combine small, dispersed water droplets into larger ones. The larger water drops are retained within a separator screen and fall to the bottom of the filter while the dry oil passes through the screen. Coalescers are capable of removing free water and solids. Vacuum dehydrator — A vacuum dehydration system removes water from oil through the application of heat and vacuum. The contaminated oil is exposed to a vacuum and heat. The water is removed as vapor. The vacuum dehydrator removes not only the free water, but also the dissolved and suspended water well below the solubility point (down to 10 ppm). In addition, vacuum dehydrators also deaerate and degasify the oil [91].

8.8.3 Refortification Refortification refers to the act of adding a predetermined amount of additive to a clean, dry, used lubricant to replenish some of the depleted additives [92]. In most cases, refortification and purification are used together.

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8.9 Fire-Resistant Fluids Fire-resistant fluids are used in the hydraulic actuators of large steam turbines operating at very high temperatures in excess of the auto-ignition temperature of turbine oil. Since the early 1970s, phosphate esters have been the only fire-resistant fluids approved by the turbine builders for use as a turbine control fluid although small amounts of more flammable carboxylate or synthetic esters have been used in refurbished systems [93].

8.9.1 Properties The main advantage of phosphate esters is their fire-resistance. Phosphate esters tend to have higher flash and fire points, higher auto-ignition temperatures and perform better in spray flammability and wick-type fire propagation tests [94]. Auto-ignition temperatures are in the region of 550–590◦ C. The triaryl version of the phosphate ester possesses inherent self-extinguishing properties because the fluid does not create enough energy to support its own combustion. Triaryl phosphates, in addition to their fireresistant properties, have good thermal stability, excellent boundary lubrication properties, low volatility, fair hydrolytic stability [94], adequate air release, and low-foaming properties. The density of phosphate esters is roughly 30% higher than mineral oils necessitating some additional consideration with respect to lube oil system design. Phosphate ester-based fluids are described in ASTM D4293. Viscosity grades are either ISO VG-32 or ISO VG-46. Phosphate ester fluids can be incompatible with some seal and insulative materials as well as certain paints thus making the pressurization system design and maintenance critical.

8.9.2 Degradation In service, phosphate esters are subject to deterioration as a result of hydrolysis, oxidation, and contamination. In the case of triaryl phosphate ester hydraulic fluids, contamination may be by water, particulates, mineral oil, and chlorine or chlorinated materials [95]. The principle degradation pathway for phosphate esters in steam turbine-generator lubrication systems is hydrolysis. While water is inevitably present in the fluids, its continued high concentration can be tolerated if fluid acidity is controlled [96]. As the solubility of water in phosphates is very much higher than in oil (reaching about 2500 ppm at 25◦ C), free water is not usually a problem and the level of fluid acidity will normally determine the suitability of the fluid for continued use. Many of the problems with the use of phosphate esters in turbine applications are associated with the development of acidity due to hydrolysis or oxidation. Since acidity development can cause corrosion, further accelerate the rate of hydrolysis, and is probably an early stage in the process of deposit formation, the maintenance of acidity levels of less than 0.5 mg KOH/g and preferably less than 0.2 mg KOH/g is strongly recommended [97]. Contamination by mineral oils can impair fire resistance, as well as being incompatible with various seals. High chlorine content can cause servo valve electrokinetic wear [95].

8.9.3 Condition Monitoring The following properties are considered necessary for the in-service testing of phosphate esters; appearance, chlorine content, color, mineral oil content, total acid number or neutralization number, fluid cleanliness, particle size, resistivity, viscosity, water content, and air release. The parameters that are of most concern are the increase in acidity, water content, and particulate contamination level. When triaryl phosphates degrade the most common result is an increase in acidity with little effect on viscosity change. Triaryl phosphate ester fluids are condemned if the acid number exceeds 0.2 over the original value (typically 0.03) [98]. Water should be kept below 2000 ppm. Alternate guidelines for maintenance and use of triaryl phosphate ester turbine control fluids can be found in IEC 60978.

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8.9.4 Maintenance The key to the cost-effective use of phosphate esters is the use of conditioning media to remove acid degradation products. Fuller’s earth and activated alumina have provided years of acceptable service; however, new adsorbents based on ion-exchange resins, may allow the fluid to be kept in the system for many years. Vacuum dehydration is required to remove the displaced water [93]. Phosphate ester hydraulic fluids require additional consideration of the lube oil system. Their use in high-pressure (1000 psi) systems requires fine filtration (0.5 to 5 µm) to protect more closely fitted pumps and control valves [99]. In addition, adsorbent filtration of phosphate ester hydraulic fluids using fullers’ earth, activated alumina, or ion exchange resin is needed to control fluid acidity. Adsorbent filters remove dissolved contaminants, such as acids, that are not removed economically or at all by other processes [100]. There is a tendency for these types of filters to remove additive materials. For this reason, adsorbent clay filters are typically not used on turbine oils, but are used for purifying fire-resistant phosphate ester hydraulic fluids as used in turbine control systems. The filters are most often used in a continuous bypass mode with 1–3% treatment ratio or they are used intermittently in accordance with changes in the acid number [97]. A fine particulate filter must be placed in series and downstream of the fuller’s earth filter to control particulates.

References [1] Steam, Its Generation and Use, 40th ed., The Babcock & Wilcox Company, New York, chap. 57. [2] Church, E.F., Steam Turbines, 3rd ed., McGraw-Hill Book Company, Inc. New York, 1950, chap. 1. [3] Electric Power Plant Design, Technical Manual TM 5-811-6, Department of the Army, Washington, DC, 1984. [4] API Standard 611, General-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [5] API Standard 612, Special-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [6] NEMA Standard SM-23, Steam Turbines for Mechanical Drive Service, National Electrical Manufacturers Association, Washington, DC, 2002. [7] NEMA Standard SM-24, Land-Based Steam Turbine Generator Sets 0 to 33,000 kW, National Electrical Manufacturers Association, Washington, DC, 2002. [8] Military Specification MIL-T-17286D, Turbines and Gears, Shipboard Propulsion and Auxiliary Steam; Packaging of, Department of the Navy, Washington, DC, 1989. [9] Military Specification MIL-T-17523, Turbine, Steam, Auxiliary (And Reduction Gear) Mechanical Drive, Department of the Navy, Washington, DC, 1982. [10] Military Specification MIL-T-17600D, Turbines, Steam, Propulsion Naval Shipboard, Department of the Navy, Washington, DC, 1982. [11] IEC Specification 60045-1, Steam Turbines — Part 1: Specifications, IEC, 1991. [12] Wills, J.G., Lubrication Fundamentals, Marcel Decker, Inc., New York, 1980, chap. 9. [13] Zeidan, F.Y. and Herbage, B.S., Fluid film bearing fundamentals and failure analysis, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 161. [14] Nicholas, J.C., Tilting pad bearing design, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 81. [15] New, N.H., Experimental comparison of flooded, directed, and inlet orifice type of lubrication for a tilting pad thrust bearing, Journal of Lubrication Technology, 96, 22, 1974. [16] Mikula, A.M., The leading edge groove tilting pad thrust bearing: recent developments, Trans. ASME, Journal of Tribology, 107, 423, 1985. [17] Gardner, W.W., Hydrodynamic Oil Film Bearings: Fundamentals, Limits and Applications, Waukesha Bearings Corporation, Waukesha, WI, 1998.

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[18] Kingsburg, G.R., Friction and wear of sliding bearing materials, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 741. [19] Ludema, K.C., Failures of sliding bearings, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 483. [20] Summers-Smith, J.D., A Tribology Casebook, Mechanical Engineering Publishing Ltd., Bury St. Edmunds, 1997, pp. 5–7, 91–92, 110–113. [21] API Standard 670, Machinery Protection Systems, 4th ed., American Petroleum Institute, Washington, DC, 2000. [22] Jackson, C., How to prevent turbomachinery thrust failures, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 152. [23] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 5B: Tilting Pad Thrust Bearings, The Glacier Metal Co. Ltd., 1989. [24] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 10: Standard Tilting Pad Journal Bearings, The Glacier Metal Co. Ltd., 1989. [25] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 14: Medium and Thick Wall Journal Bearings, The Glacier Metal Co. Ltd., 1989. [26] Snyder, D.R., Selecting rolling element bearings for modern applications: how performance criteria, operating environments, materials, lubrication requirements and other factors shape the bearingselection process, Tribology and Lubrication Technology, 60, 28, 2004. [27] Cantley, R.E., The effect of water in lubricating oil on bearing fatigure life, ASLE Transactions, 20, 244, 1977. [28] Kure-Jensen, J., Large steam turbine-generators, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 91. [29] Bloch, H.P., A Practical Guide to Steam Turbine Technology, McGraw-Hill, New York, 1996, 136. [30] Ogle, A.W. and Render, M., Impact of market changes upon power plant control, in IMechE Seminar Publication: Steam Turbine Governing and Overspeed Protection, Professional Engineering Publishing Ltd., London, 1998, 5. [31] Hayler, M.G. and Wilson, A.C.M., Lubrication of water and steam turbines, in Lubrication in Practice, 2nd ed., Robertson, W.L., Ed., Marcel Dekker, Inc., New York, 1984, chap. 4. [32] Governing Fundamentals, Manual 25195, Woodward Governing Company, 1999, chap. 7. [33] Clark, E.E., Protecting against rotating equipment loss, Chemical Engineering, 104, 106, 1997. [34] Wilson, A.C.M., Problems encountered with turbine lubricants and associated systems, Lubrication Engineering, 32, 59, 1976. [35] Neale, M.J., Drives And Seals, A Tribology Handbook, Society of Automotive Engineers, Inc., Butterworth-Heinemann, Ltd., Oxford, 1994, 29. [36] Crease, A.B., Design principles and lubrication of gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B1. [37] Calistrat, M.M., Extend gear coupling life — Part I, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 161. [38] Wright, J., Principle engineering features required by high performance gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B4. [39] Calistrat, M.M., Shaft-coupling lubrication, Lubrication Engineering, 37, 9, 1980. [40] Calistrat, M.M., Mechanical shaft couplings, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 565. [41] Calistrat, M.M., Extend gear coupling life – Part II, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 166. [42] Carter, D., Garvey, M., and Corcoran, J.P., The baffling and temperature prediction of coupling enclosures, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 115.

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[43] Errichello, R., Friction, lubrication, and wear of gears, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 541. [44] Bloch, H.P., A Practical Guide to Compressor Technology, McGraw-Hill, New York, 1996, 372. [45] Transamerica Delaval Engineering Handbook, 4th ed., Welch, H.J., Ed., McGraw-Hill Book Company, New York, 1983, chap. 5. [46] Low, M.B.J., The sliding of high temperature components in steam turbines, Proceedings of IMechE, 194, 171, 1980. [47] Hackel, R.A. and Keyes, H.M., Steam turbines in process industries, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 4. [48] Steam turbines and auxiliaries, Power, 133, June, 1989. [49] Pilicy, F.X. and Dundas, R.E., Insurance for steam turbines, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 10. [50] Westhofen, B., Enhancing the availability of industrial turbines, Brown Boveri Review, 73, 31, 1986. [51] Cameron, J.A., Materials for use in geothermal steam turbines, in Geothermal ’77, Carrier Corporation, 1977, 39. [52] Ortolano, R.J., Steam turbine blading maintenance: part IV, Turbomachinery International, 24, 56, 1983. [53] D’Innocenzio, M., Oil systems – design for reliability, Proceedings of the First Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1970, 163. [54] Ventilation of steam turbines, Gargoyle Lubricants Technical Bulletin, 1945. [55] Sassos, M.J., Babyak, M.R., and Zerbe, J.P., Lubrication and seal oil systems — common problems and practical solutions, Proceedings of the 22nd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1993, 51. [56] Salisbury, R.J., Stack, R., and Sassos, M.J., Lubrication and seal oil systems, Proceedings of the 13th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1984, 151. [57] API Standard 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1999. [58] API Recommended Practice 686, Recommended Practice for Machinery Installation and Installation Design, 1st ed., American Petroleum Institute, Washington, DC, 1996. [59] Petroleum Products, Lubricants, and Fossil Fuels, Annual Book of ASTM Standards, Section 5, ASTM International, West Conshochocken, PA, 2003. [60] Schwager, B.P., Hardy, B.J., and Aguilar, G.A., Improved response of turbine oils based on Group II hydrocracked base oils compared with those based on solvent refined base oils, in Turbine Lubrication in the 21st Century, ASTM STP 1407, Herguth, W.R. and Warne, T.M., Eds, American Society for Testing and Materials, West Conshohocken, PA, 2001, 71. [61] Hannon, J.B., How to select and service a turbine oil, Machinery Lubrication, July, 2001. [62] Potential performance problems caused by mixing different types of turbine oils, Engineering and Construction Bulletin 2003-17, US Army Corps of Engineers, 2003. [63] Dolby, G.W. and Kofke, W.A., The role of base oils and additives in modern-day turbine oils, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 1. [64] O’Brien, J.A., Lubricating oil additives, in CRC Handbook of Lubrication (Theory and Practice of Tribology), Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1984, 301. [65] von Fuchs, G.H., Rust inhibitors — their evaluation and performance, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 28. [66] vd Merwe, D.G., Morgan, P.M., Roets, P.N.J., and Botha, J.J., Additives in turbine and dynamic compressor oils, The South African Mechanical Engineer, 47, 13, 1997.

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[67] Swift, S.T., Butler, K.D., and Dewald, W., Turbine oil quality and field application requirements, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, American Society for Testing and Materials, West Conshochocken, PA, 2001, 25. [68] Roberton, R.S., Background and development of ASTM D 4378: practice for in-service monitoring of mineral turbine oils for steam and gas turbines, in Turbine Oil Monitoring, ASTM STP 1021, Young, W.D. and Roberton, R.S., Eds, American Society for Testing and Materials, Philadelphia, PA, 1989, 3. [69] The Lubrication of Steam Turbines, Shell, Alabaster Passmore Sons, Ltd., London, 1958. [70] Summers-Smith, D., The unacceptable face of lubricating oil additives, Tribology, 11, 318, 1978. [71] Engineering and Design — Lubricants and Hydraulic Fluids, Engineer Manual EM 1110-2-1424, Department of the Army, Washington, DC, 1999, chap. 12. [72] Furby, N.W., Hanly, F.J., and Vincent, J.A., Rusting in turbine oil systems, in Symposium on Steam Turbine Oils, ASTM STP 211, American Society for Testing and Materials, Philadelphia, PA, 1956, 40. [73] Layne, R.P., Vapor space corrosion inhibition of steam turbine lubricating and cleaning oils, in Turbine Lubricating Problems, ASTM STP 437, American Society for Testing and Materials, Philadelphia, PA, 1968, 73. [74] Steam Turbines and their Lubrication, Mobil Oil Corporation, New York, NY, 1981. [75] Enz, W.E. and Hausermann, A., Particular problems of steam turbine lubrication, Proceedings of the Seventh Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1981, 125. [76] Steam Turbine Lubrication, 2nd ed., The Texas Company, 1947. [77] Fowle, T.I., Problems in the lubrication systems of turbomachinery, Proceedings of the Instrumental Mechanical Engineers, 186, 705, 1972. [78] Bice, C.D., Air entrainment issues in equipment using new generation turbine oils, Presented at STLE, Pittsburgh Section, November, 2004. [79] Fitch, J.C. and Jaggernauth, S., Moisture — the second most destructive lubricant contaminate, and its effects on bearing life, P/PM Magazine, December, 1994. [80] Coleman, W.L., Water contamination of steam turbine lube oils — how to avoid it, Proceedings of the Seventeenth Turbomachinery Symposium, 51. [81] Missana, A. and Steenburgh, J.H., Ensuring clean lube oil for large steam turbines, Power Engineering, 88, 46, June, 1984. [82] Bissett, W., Cost effective condition monitoring of large steam turbine/generator oil systems, Transactions of Mechanical Engineering, IEAust., ME20, 61–68, 1995. [83] Abner, Jr., E., Lubricant deterioration in service, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 517. [84] Lamping, G.A., Cuellar, Jr., J.P., and Silvus, H.S., Summary of maintenance practices for large steam turbine-generator lubrication systems, ASME/IEEE Power Generation Conference, ASME Paper 86-JPGC-Pwr-14, 1986. [85] DenHerder, M.J. and Vienna, P.C., Control of turbine oil degradation during use, Lubrication Engineering, 37, 67, 1981. [86] McCloskey, T.H., Troubleshooting bearing and lube oil system problems, Proceedings of the 24th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1998, 147. [87] Ohgake, Ryoji, Sunami, M., Yoshida, T., and Watanabe, J., The reliable control of oil quality in Japanese turbine units, in Turbine Oil Monitoring, Young, W.C. and Roberton, R.S., Eds, ASTM STP 1021, American Society for Testing and Materials, Philadelphia, PA, 1989, 35. [88] Bloch, H.P., Criteria for water removal from mechanical drive steam turbine lube oils, Lubrication Engineering, 36, 699, 1980.

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[89] Roberton, R.S., ASTM in-service monitoring program for steam and gas turbine oils, Lubrication Engineering, 42, 466, 1986. [90] Micetic, J., New Generation of turbine oils, US Army Corps of Engineers, Hydroelectric Design Center, Portland District. [91] Adams, M.A. and Bloch, H.P., Vacuum distillation methods for lube oils increase turbomachinery reliability, Proceedings of the 17th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1991, 41. [92] Stein, W.H. and Bowden, R.W., Shell Global Solutions, Inc., Turbine oil reclamation and refortification, Machinery Lubrication Magazine, July, 2004. [93] Phillips, W.D., The use of a fire-resistant turbine lubricant: Europe looks to the future, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, 2001, 1. [94] Brown, K., Utility Service Associates, Condition-monitoring of phosphate ester hydraulic fluids, Machinery Lubrication Magazine, November, 2002. [95] Brown, K.J. and Staniewski, J.W., Condition monitoring and maintenance of steam turbinegenerator fire resistant triaryl phosphate control fluids, in Condition Monitoring and Preventive Proceedings, STLE SP-27, Society of Tribologists and Lubrication Engineers, 1989, 91. [96] Wolfe, G.F., Experience with phosphate ester fluids as industrial steam turbine-generator lubricants, Lubrication Engineering, 8, 413, 1978. [97] Phillips, W.D., The conditioning of phosphate-ester fluids in turbine applications, Lubrication Engineering, 12, 766, 1983. [98] Thibault, R., Converting to condition-based oil changes — part II, Practicing Oil Analysis Magazine, March, 2001. [99] Steele, F.M., Filtration and reclamation of turbine oils, Lubrication Engineering, 34, 252, 1978. [100] Steele, F.M., Contamination control in turbomachinery oil systems, Lubrication Engineering, 40, 487, 1984.

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9 Compressors and Vacuum Pumps 9.1 9.2

Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Part I: Basic Principles of Gas Compression . . . . . . . . .

9.3

Part II: Compressor Design and Lubrication . . . . . . . .

9-1 9-2

Gas Laws • Gas Compression Cycle

T. Kazama Muroran Institute of Technology Department of Mechanical Systems Engineering

G.E. Totten Portland State University Department of Mechanical and Materials Engineering

9-7

Compressor Classification • Compressor Lubrication • Reciprocating Positive Displacement Compressors • Rotary Positive Displacement Compressors • Dynamic Compressors • Lubricants for Compressors

9.4

Part III: Vacuum Pump Design and Operation . . . . . .

9-45

Introduction to Vacuum Pumps • Fundamental Vacuum System Relationships • Vacuum Pump Selection Criteria • Vacuum Pumps and Applications • Seals for Vacuum Pumps • Vacuum Measurement • Lubricants for Vacuum Pumps

References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

9-57

9.1 Introduction Compressors and vacuum pumps are vitally important industrial equipment. Gas compressors are mechanical devices that are used to pressurize and circulate gas through a process, facilitate chemical reactions, provide inert gas for safety or control systems, recover and recompress process gas, and maintain correct pressure levels by adding and removing either gas or vapors from a process system. Compressor and hydraulic pumps are similar devices with the main difference being that the fluid delivered by compressors is usually a gas which is compressed and under pressure at the time it is delivered, even if there is no load on the system. Gas compressors are used in nearly every industry sector including: automotive, steel, mining, food, gas and petroleum production, and storage and energy conversion. Global demand for compressor pumps is expected to grow and additional market opportunities exist for compressor manufacturers in the global energy production sector due to increasing consumption. Similar equipment market opportunities are expected in the natural gas extraction and transportation industries. Compressors are used not only for gas compression but also for cooling air conditioning and refrigeration. Refrigeration is that process used to remove heat from a material or gas so that its temperature is less than that of its surroundings. A refrigerant is the working fluid used to transmit heat in a refrigeration compressor, which is used to remove heat from the heat-laden refrigerant vapor in the evaporator

9-1

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and which compresses the refrigerant gas to a pressure that will liquefy in the condenser. Refrigeration compressors are either electrically or mechanically driven and they are used in stationary applications such as home and business air-conditioning and in mobile applications which include automotive and truck air conditioning and refrigeration. Rapid and significant market growth is expected for most types of refrigeration compressors. Compressors are loosely related to vacuum pumps. A compressor is used to reduce the volume of a gas, whereas a vacuum pump is essentially a compressor which operates with an intake pressure below atmospheric pressure. Vacuum pumps are vitally important equipment for nearly every industrial sector including the electronic semiconductor industry, steel making, automotive manufacturing, forestry, aerospace, and many others. Compressor manufacture is not only an enormous and growing industry, nearly all compressors require a lubricant to assure adequate cooling, sealing, and lubrication of internal components. The choice of a compressor lubricant depends on the type and construction of the compressor, the gas being compressed, the degree of compression, and the final outlet temperature. Therefore, it is important to understand the operation of various types of compressors and their lubrication mechanisms in order to select a proper lubricant for the compressor. This chapter contains three parts. Part I presents basic equations relating to gas compression. Part II provides an overview of industrial gas and refrigeration compressor design and lubrication requirements. Part III provides an overview of vacuum pump design and operation.

9.2 Part I: Basic Principles of Gas Compression 9.2.1 Gas Laws For the purpose of this discussion, it will be assumed that gases being compressed by a compressor follow the well-known Ideal Gas Laws. For example, Boyles Law states that pressure times volume of a gas is a constant if the temperature is constant: p1 V1 = p2 V2 The reference value used to determine pressure must be indicated. If the reference value is a vacuum, then the pressure is absolute pa . However, if the reference value is atmospheric pressure (patm ) then it is called gauge (pg ) pressure. They are related by: pa = pg + patm Absolute pressure must be used for ideal gas law relationships. When the pressure is constant, Charles Law states that the volume of a gas increases proportionately with temperature: V1 /V2 = T1 /T2 If the temperature of a gas increases as the pressure increases and the volume is kept constant, then Amonton’s Law is followed: p1 /p2 = T1 /T2 For these calculations, all temperatures are in reference to absolute zero. If the temperature is in ◦ C, then degrees Kelvin is calculated from: Kelvin = ◦ C + 273 Similarly, if the temperature is given in ◦ F, then degrees Rankine is calculated from: Rankine = ◦ F + 460

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Compressors and Vacuum Pumps

9-3

The well-known Ideal Gas Law is obtained by combining the expressions for Boyle’s and Charle’s Laws: p1 V1 /T1 = p2 V2 /T2 Avogadro’s Law states that equal volumes of gas at the same temperature and pressure will contain the same number of molecules: pV = nRT where n is the number of moles, R is the gas constant that is selected to be consistent with the units of temperature, pressure, and volume used in the calculation. Dalton’s Law states that the total pressure (pt ) of a mixture is equal to the sum of partial pressures of each component (a, b, c, etc.): pt = pa + pb + pc + · · · . Similarly, the total volume is equal to the sum of the partial volumes of the constituent gases according to Amagat’s Law: V T = Va + V b + V c + · · · . If the temperature of a gas decreases or if the pressure increases sufficiently, the gas will undergo a change of state to a liquid. Further decreases in temperature or increases in pressure will convert the liquid into a solid. If the temperature increases sufficiently, a point will be achieved where the gas can no longer be liquefied by increasing pressure. The highest temperature at which a gas can be liquefied by increasing pressure is called the critical temperature of the gas. The pressure required to liquefy the gas at the critical temperature is called the critical pressure. When the pressure, temperature, and volume variation of a gas follow the Ideal Gas Law, the gas is referred to as an ideal gas. However, as the pressure increases, the behavior of a gas deviates from that predicted by the Ideal Gas Law. This is due to the compressibility of the gas and is accounted for in the Ideal Gas Law calculation by using the compressibility factor (Z ): p1 V1 /T1 Z1 = p2 V2 /T2 Z2 Values for compressibility factors for gases are obtained from charts called “general compressibility charts,” which are typically published in reference books. A series of thermophysical constants for commonly encountered compressed gases is provided in Table 9.1.

9.2.2 Gas Compression Cycle Figure 9.1 shows the p–V curve, namely, the relation of volume V , pressure p, and work [1]. Work is equal to force × distance where pressure corresponds to the force on the cylinder and volume corresponds to the distance the compressor piston moves. The area under the curve p × V is equal to the work performed during the gas compression cycle. There are two ways that a positive displacement compressor can be operated: either isothermally or adiabatically. For isothermal operation, temperature is held constant during compression by removal of the heat of compression and the work corresponds to: p1 V1 = p2 V2 Adiabatic compression (or expansion) occurs when there is no heat added or removed during the process. In adiabatic processes, pressure will vary with an exponential value of volume: p1 V1k = p2 V2k

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9-4

TABLE 9.1

Thermophysical Constants of Selected Gases

Gas

a

NH3 CO2 H2 N2 O2 H2 S CH4 C 2 H6 C 2 H4 C 3 H8 C 3 H6 C4 H10 C4 H10

Molecular weight (g/mol)

Critical temperature (◦ C)

Critical pressure (bar)

28.95 17.03 44.01 2.016 28.0134 31.9988 34.08 16.043 30.069 28.054 44.096 42.08 58.123 58.123

−140.5 132.4 31 −240 −147 −118.6 100 −82.7 32.2 9.5 96.6 91 152 134.9

37.71 112.8 73.825 12.98 33.999 50.43 89.37 45.96 48.839 50.76 42.5 46.1 37.96 36.48

1.202 0.73 1.87 0.085 1.185 1.354 1.45 0.68 1.282 1.178 1.91 1.81 2.52 2.51

Specific heat at CP CV (1.013 bar (1.013 bar and 21◦ C and 21◦ C [70◦ F]) [70◦ F]) kJ/(mol K) kJ/(mol K) 0.029 0.037 0.037 0.029 0.029 0.029 0.034 0.035 0.053 0.042 0.075 0.062 0.097 0.095

0.02 0.028 0.028 0.021 0.02 0.021 0.026 0.027 0.044 0.034 0.066 0.054 0.088 0.086

CP /CV Gamma-value

Specific volume (1.013 bar and 21◦ C [70◦ F]) m3 /kg

Compressibility (Z -factor) (1.013 bar and 15◦ C [59◦ F])

1.4028 1.309623 1.293759 1.384259 1.403846 1.393365 1.326 1.305454 1.193258 1.242623 1.134441 1.156832 1.093586 1.095845

0.833 1.411 0.547 11.986 0.862 0.755 0.699 1.48 0.799 0.862 0.543 0.587 0.4 0.406

0.9992 0.9929 0.9942 1.001 0.9997 0.9994 0.9915 0.998 0.9912 0.9935 1.0193 0.984 0.9625 0.9675

a The composition of air is: N = 78.09% by vol.; O = 20.94% by vol.; Ar = 0.93% by vol.; CO = 330 ppm by vol.; Ne = 18 ppm by vol.; He = 5.2 ppm by vol.; 2 2 2 Kr = 1.1 ppm by vol.; H2 = 500 ppb by vol.; Xe = 86 ppb by vol.; and Rn = 6 × 10−11 ppb by vol. b The data was obtained from Air Liquide from their Web site: http://www.airliquide.com/en/business/products/gases/gasdata/list.asp. c Except where otherwise noted, the data was obtained from Air Liquide website address provided in Note b.

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Handbook of Lubrication and Tribology

Airb Ammoniab Carbon dioxideb Hydrogenb Nitrogenb Oxygenb Hydrogen sulfidec Methaneb Ethaneb Ethyleneb n-Propaneb Propyleneb n-Butaneb i-Butaneb

Chemical formula

Density at (1.013 bar and 15◦ C [59◦ F]) kg/m3

Compressors and Vacuum Pumps

9-5

FIGURE 9.1 p–V curve. (From Figure 4 in Garg, D., Totten, G.E., and Webster, G.M., Compressor Lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. G.E. Totten, ASTM International, USA, 2003, chap. 14. With permission.)

where k is the ratio of specific heats. When the adiabatic process is reversible, it is referred to as an isotropic process (k = 1). When the operation is assumed to be adiabatic, the discharge temperature T2 can be easily estimated. For example, if T1 = 293 K, p2 /p1 = 11 and n = 1.4 (the ratio of specific heats, Cp /CV = k, of air), T2 is determined to be 308◦ C. This example indicates that the greater the amount of gas compression, the higher the final temperature. Although less work is required for an isothermal process, it is impossible to achieve it even though compressors are designed for as much heat removal as possible. Similarly, adiabatic processes are also impossible to achieve since some heat is always added or emitted in use. Therefore, what is typically achieved is a polytropic cycle: p1 V1n = p2 V2n where n (the polytropic index) is experimentally determined for each type of compressor and is usually not equal to k. The index n is unity for isothermal operation and n is the specific heat of the gas for adiabatic operation. (Thermodynamically, polytropic compression is defined as that process where: pV n = constant. A polytropic process differs from an adiabatic process in that the change does not occur at constant entropy since heat is either added to, or removed from, the gas. When heat is extracted by cooling, the n-value will be less than the adiabatic k-value.) The value of n may also be calculated from: T2 /T1 = (p2 /p1 )(n−1)/n where p is the absolute pressure, T is the absolute temperature, and the subscripts 1 and 2 are corresponding to the suction and discharge period respectively. Thermodynamically, adiabatic and isothermal processes are reversible but polytropic processes are irreversible, steady-state processes. A typical gas compression cycle where a gas is compressed in a piston cylinder from the inlet pressure ps to the discharge pressure pd along the line 1-2 is shown in Figure 9.2 [1]. Since it is impossible to discharge all of the gas due to the volume of the space not covered by the piston stroke, there will be a residual

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Handbook of Lubrication and Tribology

FIGURE 9.2 p–V cycle and piston stroke. (From Figure 3 in Garg, D., Totten, G.E., and Webster, G.M., Compressor Lubricants, in Fuels and Lubricants Handbook: Technology, Properties, Performance, and Testing, ed. Totten, G.E., ASTM International, USA, 2003, chap. 14. With permission.)

amount of gas referred to as the clearance volume. This is typically the area between the cylinder and the head of the piston illustrated in Figure 9.2(a). Clearance volumes typically range from 4 to 20%. Figure 9.2(b) illustrates the completion of the compression stroke along path 2-3. When the piston reaches the point 3, the discharge valve closes and the piston undergoes the expansion stroke 3-4 (shown in Figure 9.2[c]) until the pressure drops below the inlet pressure at the point 4. At the point 4, the inlet valve opens and the gas fills the cylinder as shown in Figure 9.2(d) and the process is repeated. Figure 9.2 illustrates the relation of the p–V curve and the piston stroke [1]. The schematic representations illustrated by Figures 9.2(a), (b), (c), and (d) correspond to the compression, discharge, expansion, and suction strokes respectively. There is a residual value referred to as clearance (dead) volume, since all of the gas is unable to discharge. Typical values of dead volumes range from 4 to 20%. The piston stroke is also illustrated in Figure 9.2. For Figures 9.2(c) and (d), the actual stroke is less than that represented by the line 4-1. The ratio of the actual capacity to the total displacement is referred to as volumetric efficiency. The volumetric efficiency is always less than the theoretical value because: • • • •

The reexpansion of the gas trapped in the cylinder clearances Entrance losses due to the pressure drop at the inlet Piston valve and ring leakage Slight increase in gas volume due to the heat rise from the warm cylinder

Theoretical volumetric efficiency ηv for polytropic compression for a diaphragm compressor is calculated from: ηv1 = 100 − C(R 1/k − 1) = 100 + C − CR 1/n

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Compressors and Vacuum Pumps

9-7

where R is the compression ratio that is defined as absolute discharge pressure divided by the absolute inlet pressure of the compressor and C is the cylinder clearance (%). For a piston pump, the equation for volumetric efficiency is: ηv2 = 100 − C(R 1/k − 1) = 100 + C − CR 1/k where the symbol  is a leakage factor for piston compressors. It is determined experimentally for compression ratios of 6 to 10:1. The above equations refer to ideal gases. To be more generally applicable, these equations must be modified to account for the influence of the compressibility of the gas. This is done by using the compressibility factors for the gas at the inlet (Z1 ) and outlet (Z2 ) resulting in the following corrected equations for a diaphragm: ηvZ1 = (100 − C(R 1/k − 1) = 100 + C − CR 1/n )(Z1 /Z2 ) ηvZ2 = (100 − C(R 1/k − 1) = 100 + C − CR 1/k )(Z1 /Z2 ) For an intermediate pressure air compressor using a petroleum lubricant, the correction of the volumetric efficiency may be approximately 5%. It is necessary to determine the amount of power required to drive the system in order to properly size a compressor. This is done by determining the amount of brake horsepower required to compress a given volume of gas from the incoming inlet pressure to the desired discharge pressure [1]. Brake horsepower is defined as the ideal isoentropic (theoretical) horsepower plus any fluid (valve, flow, and other leakage) or mechanical friction losses. Theoretical horsepower (hptheory ) may be calculated from [1]: hptheory =

144 33,000




   k (r − 1)k−1/k ps Vs k −1

where the units of the value 144 is in.2 /ft2 , 33,000 foot-pounds/min = 1 hp, k = the k-value for the gas, ps is the inlet pressure in PSIA, Vs is the inlet volume in ft3 /min, and R is the compression ratio. Compressors may also be characterized by specific power consumption pspec [1]: pspec = power consumption (kW)/volume flow (m3 /min) As indicated by the gas laws described earlier, gas compression is accompanied by a temperature rise. The greater the compression, the greater the temperature rise. If high discharge pressures are required, the compression process must be accompanied by two or more cooling stages, which improves efficiency and reduces power consumption. The temperatures actually encountered in gas compression must be considered in lubricant selection since the viscosity of the lubricant is temperature dependent and increased temperatures increase the potential for oxidation and deposit formation.

9.3 Part II: Compressor Design and Lubrication There are numerous compressor designs available. In this section, the design and operation of only the more commonly encountered compressor types will be discussed.

9.3.1 Compressor Classification Compressors, blowers, and fans increase the pressure of a gas or mixture of gases such as air, nitrogen, or a refrigerant and move it to where it can be applied to move actuators or refrigerating cycles [2,3]. Compressors may be classified in terms of their discharge pressure: high (>2000 kPa, gauge), intermediate

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Handbook of Lubrication and Tribology

(800–2000 kPa, gauge), and low (100–800 kPa, gauge). Compressors should be contrasted with blowers that are used to handle large volume of gases at gauge pressures ranging from 10 to 100 kPa and fans that are used at gauge pressures, which are typically less than 10 kPa [4]. Only compressors and their lubrication will be addressed here. Typical ISO 6743 compressor categories are provided in Table 9.2. There are two functional types of compressors: positive displacement and dynamic (or turbo). Displacement compressors increase the pressure of the gas by reducing its volume [5–7]. This is accomplished by taking in successive volumes of air that is confined within a closed space (such as a piston in a cylinder) and elevating the entrapped gas to a higher pressure. Displacement compressors are preferred for high pressures and relatively small volumes of gas. Examples of displacement compressors include: rotary and reciprocating. Rotary pumps can be one-rotor such as: sliding vane, liquid piston (liquid ring), rolling piston, scroll, and single screw types or they may be two-rotor types such as lobe or screw compressors. Reciprocating compressors include: crosshead, trunk, labyrinth (reciprocating piston), diaphragm compressors, and rocking piston compressors. Reciprocating compressors are available either as aircooled or water-cooled in lubricated and nonlubricated configurations, may be packaged, and provide a wide range of pressure and capacity selections. Dynamic compressors are turbo-compressors where impellers transfer rotational energy to a gas to be compressed. Examples of dynamic (or turbo) compressors include axial, radial, and centrifugal compressors. Centrifugal compressors, for example, produce high-pressure discharge by converting angular momentum imparted by the rotating impeller (dynamic displacement). To do this efficiently, centrifugal compressors rotate at speeds that are higher than those of other types of compressors. Dynamic compressors are also designed for higher capacity because flow through the compressor is continuous. Dynamic compressors may be further subclassified by: the number of compression stages, cooling method (air, water, oil), drive method (motor, engine, steam, other), lubrication (oil, oil-free), packaged or custombuilt. However, with the exception of lubrication, these additional classifications will not be discussed in detail here.

TABLE 9.2 Category of Compressors Compressors Positive displacement Reciprocating Piston Single acting Double acting Axial piston Swash plate Rotating Straight lobe Roots Multi-lobe Rotary Sliding vane Rolling piston Swing Screw Single Twin Scroll Liquid piston (Liquid ring) Dynamic (Turbo) Centrifugal (Radial) Axial Mixed

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Compressors and Vacuum Pumps TABLE 9.3

9-9

Duty Classification for Reciprocating Oil-Lubricated Air Compressors Operating conditions

Duty

Symbol

Normald

DAA

Severee

DAB

a b c d e

Intermittent or continuous operation Intermittent or continuous operation

Discharge temperaturea (◦ C)

Differential pressureb (bar)

Discharge pressurec (bar)

2500 psi), this type of pump is usually thought of as a low-pressure pump (2900

TABLE 10.8

Equivalent Length Values Equivalent length (Le /D)

Device Check valve 90◦ Standard elbow 45◦ Standard elbow Close return bend Standard tee-run Standard tee-branch

150 30 16 50 20 60

the following equation: Nr =

3162 × Q µd

(10.47)

where Nr is Reynolds number, µ is viscosity (cSt), and d is pipe I.D. (in.). Fittings and valves must be handled somewhat differently than straight runs of pipe. The easiest way to calculate the losses due to fittings and valves is to use the equivalent length method to estimate the effect by treating it as if it were an additional length of pipe. Table 10.8 lists some common devices and their equivalent length values, which are given as the length-to-diameter (Le /D) ratios so that they can be used directly in the modification of the Darcy equation as follows: hr = λ ×

Le v2 × D 2g

(10.48)

where hf is equivalent length, λ is friction factor, Le /D is equivalent length values, v is fluid velocity, and g is gravitational constant. The analytical methods presented here to calculate pressure losses in hydraulic piping and fittings are accurate but can be very time consuming. A method that is less accurate but provides a reasonable estimate of pressure losses in hydraulic systems involves the use of tables, which are available from pipe manufacturers and in various handbooks concerning fluid flow (Table 10.5). 10.2.4.2 Reservoir Design A typical design for an industrial reservoir is shown in Figure 10.54 [21]. Several features can be seen in this figure. The overall dimensions should enclose a sufficient volume of oil to permit air bubbles and foam to escape during the resident time of the fluid in the reservoir. The depth must be adequate to assure that during peak pump demands the oil level will not drop below the pump inlet level. The pump should be mounted below the reservoir so that a positive head pressure is available at all times. This is very critical when water based hydraulic fluids are used, since these fluids can have a higher specific gravity as well as a

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10-44

Handbook of Lubrication and Tribology Mounting plate Suction line Filler / Breather cap

Oil level gauge

Dished bottom

Cleanout cover

Return line Drain line

Drain plug

Baffle plate

FIGURE 10.54 A typical design for an industrial reservoir.

much higher vapor pressure than mineral oil-based fluids (Section 10.2.2.7). The reservoir should be sized to afford adequate fluid cooling. Baffles are provided to prevent channeling of the fluid from the return line to the inlet line. The bottom of the return line is usually cut at a 45◦ angle to assist in redirection of the fluid away from the inlet. A clean-out plate is provide to promote cleaning and inspection. Sight gages are normally used to monitor the fluid level. A breather system with a filter is provided to admit clean air and to maintain atmospheric pressure as fluid is pumped into and out of the reservoir. With water-based hydraulic fluids a pressurized reservoir is recommended. Special breather caps can be purchased to vent between 1 and 15 psig. If one of these is used make sure that it has a vacuum brake to vent at ∼−0.5 psig (Note: Not all pressure caps have a vacuum brake). This is important so that when the reservoir is cooling down no appreciable vacuum develops in the reservoir tank. This feature will minimize pump cavitation upon start-up and also prevent a possible tank implosion. 10.2.4.3 Natural Frequency and Time Response When designing any hydraulic system, especially when heavy masses are moved quickly, there is one very important design factor that needs to be considered. That factor is known as the “natural frequency” (ωo ) of the system. Knowledge of this frequency is important because it determines how fast one can accelerate a given load and thus its maximum achievable velocity. From the physical laws of motion, the natural frequency of a hydraulic system can be found by taking the square root of the effective spring constant divided by the effective moving mass.
ωo =

C M

(10.49)

where ωo is natural frequency, C is effective spring constant, and M is effective moving mass. This is a simple statement; however, determination of the effective spring constant and effective moving mass is not so simple. The effective spring constant not only includes the compressibility of the trapped hydraulic fluid between the valves and the actuators, but also the movement of any hoses or piping as well as structural vibrations. The effective mass of the system is the combination of all the moving loads, including the mass of the trapped fluid between the valves and actuators as illustrated in Figure 10.55 [23].

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Basic Hydraulic Pump and Circuit Design

10-45

M1

M2 Mechanical component

Hydraulic spring-mass system

Valve

FIGURE 10.55 Illustration of the hydraulic spring-mass system.

Ae Ab V1

M V2

FIGURE 10.56 Illustration of the natural frequency parameters.

In the simplified case of a linear cylinder in a closed circuit (Figure 10.56) [23], the natural frequency can be calculated using the following expression:  ωo =

Ab2 × β A2 × β + e V1 × M V2 × M

or

Fo =

ωo 2π

(10.50)

where ωo is natural frequency (rad/sec), Ab is cylinder blind end area (in.2 ), Ae is cylinder extending end area (in.2 ), β is bulk modulus of fluid, V1 is cylinder blind end volume (in.3 ), V2 is cylinder extending end volume (in.3 ), M is effective moving mass (lbs-sec2 /ft, slugs), and Fo is natural frequency (Hz). There are computer programs available that can be used to determine the frequency response of a hydraulic system by using the impulse method. The time response of a hydraulic system is the synergistic result of the response times of all of the components used in the system [24]. Therefore, most component manufacturers will provide information relative to the responsiveness of their components. Unfortunately, the information derived from the component manufacturers is not consistent. The ability to understand and utilize the response information obtained from component manufacturers using a second-order system depends upon the definition of several aspects of the response subject as follows: • Delay time: the time required for the output to reach 50% of the steady output. • Rise time: the time required for the output to rise from 10% to 90% of the final output value. • Maximum overshoot: the time at which the maximum overshoot occurs. • Settling time: the time for the system to reach and stay within a stated plus-and-minus tolerance band around the steady-state output.

© 2006 by Taylor & Francis Group, LLC

10-46

Handbook of Lubrication and Tribology Maximum overshoot time

c(t) 1.5

Unit step input

Tolerance band 6d

1+d 1.0 1–d 0.90

Steady-state error (t → `)

0.5 Delay time

0.1 0.0 Rise time Settling time

FIGURE 10.57

Time

Step response of a second-order system.

S

(S – d )

d

Ae Ab L 1 V1

FIGURE 10.58

M V2 L 2

Illustration of a typical natural frequency calculation.

A graph, which illustrates these parameters, is shown in Figure 10.57. Control technology can be used to evaluate the response of a complete hydraulic system if all of the component information is given in consistent and correct terms. 10.2.4.3.1 Calculation of Natural Frequency, Acceleration, Maximum Velocity, Acceleration Pressure, and Flow Rate For economic reasons, it is often desirable to operate a hydraulic system as fast as possible. This is especially true on automated assembly lines where hydraulics is used to move parts. As an example of a simple calculation, consider the following application where one needs to determine the maximum speed and shortest cycle time to perform a repetitive task. By way of a single-rod hydraulic cylinder (1.5

bore, 1

rod), a proportional directional control valve is used to accelerate a 1000 lb load (M ) to a constant velocity over a distance of 30 in. in 1 sec and then decelerate the load to a stop. The load is then retracted in the same manner to start the cycle over again (Figure 10.58) [23]. To solve this problem, the natural frequency must first be calculated so that the time to accelerate the load can be determined.

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Basic Hydraulic Pump and Circuit Design

10-47

Then the maximum velocity, acceleration pressures, and required flow rates can be calculated for both the extending and retracting modes. The following information is given: w = 1000 lbs. (Load) Ts = 1.0 sec (Stroke time) Ab = 1.76 in.2 (1.5

Cylinder bore) Ae = 0.98 in.2 (1.5

bore area − 1

rod area) S = 30 in. (Stroke distance) β = 200,000 lb/in.2 (Bulk Modulus of oil) L1 = 46.50 in. (Cylinder blind-end pipe length) L2 = 38.75 in. (Cylinder rod-end pipe length) D = 0.62 in. (Pipe I.D.)

Pipe Size = 3/4

O.D. × 0.065

wall The first step is to calculate the pipe-trapped volumes between the control valve and the cylinder blind-end inlet (V3 ) and the rod-end inlet (V4 ) in Figure 10.58 [23] from the following equation: π D2 × L1 4 π(0.62)2 = × 46.5 4

V3 =

= 14.04 in.3

(10.51)

V4 = 0.30 × 38.75 = 11.7 in.3 Next, calculate the dimension “d,” using the above values along with the given parameters, using the following expression:   [(Ae × S + V4 )/ Ae3 ] − (V3 / Ab3 ) d= √ √ (1/ Ae ) + (1/ Ab ) √ √ {[(0.98 × 30) + 11.7] / 0.983 } − (14.04/ 1.763 ) = √ √ (1/ 0.98) + (1/ 1.76) = 20.6 in.

(10.52)

Next, calculate the total trapped volume between the valve and cylinder blind-end (V1 ) and cylinder rod-end (V2 ), using the following relations: V1 = V3 + (Ab × d) = 14.04 + 1.76 × 20.6 = 50.3 in.3 V2 = V4 + Ae (S − d) = 11.7 + 0.98(30 − 20.6) = 20.9 in.3

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(10.53)

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Handbook of Lubrication and Tribology

Convert the load to units of mass as follows: M=

1000 w = = 2.59 (lbs-sec2 /in.) g 386

(10.54)

where M is effective moving mass (lbs-sec2 /in., slugs), w is load force (lbs), and g is gravitational constant (in./sec2 ). Then substitute the known quantities into Equation (10.50) to obtain the natural frequency (ωo ) of this system as follows:  ωo =

1.762 × 200,000 0.982 × 200,000 + 50.3 × 2.59 20.9 × 2.59

= 91.1 rad/sec

(10.55)

In calculating “ωo ” we have not taken into consideration other factors which contribute to the spring constant of the system, namely, hoses and other mechanical components. However, it has been shown over the years, that a good approximation to determine the useable acceleration is to divide the calculated natural frequency by three [23]. This simplification avoids a much more complex mathematical analysis, which would have required variables, which are difficult, if not impossible to define. Therefore, the useable frequency (ω) can be estimated as follows: ω=

ωo 91.1 = = 30.4 rad/sec 3 3 or

F=

(10.56)

30.4 ω = = 4.8 Hz 2π 2π

The acceleration time (T ) or the time for one complete oscillation can now be calculated: T =

1 1 = = 0.033 sec ω 30.4

(10.57)

However, it has been determined that this period is too short for acceleration to stabilize using proportional valves. Generally, for stable acceleration, the time allowed must be a minimum of four to six times the time period for one oscillation [23]. Therefore, the acceleration stabilizing time (Tb ) is calculated as: Tb = 6 × T = 6 × 0.033 = 0.20 sec

(10.58)

From the stroke distance (S), the acceleration time (Tb ) and the stroke time (Ts ); the maximum velocity (Vmax ), acceleration (Amax ), and the acceleration force (Fa ) can be easily calculated from the following expressions: S 30 = = 37.5 in./sec Ts − T b 1.0 − 0.2 Vmax 37.5 = 188 in./sec2 = Amax = Tb 0.20 w 1000 × 188 = 487 lbs Fa = MA max = Amax = g 386 Vmax =

© 2006 by Taylor & Francis Group, LLC

(10.59) (10.60) (10.61)

Basic Hydraulic Pump and Circuit Design

10-49

Before we can calculate the acceleration pressure at the blind-end (Pb ) and rod-end (Pr ) of the cylinder, the frictional force that the load imposes on the system needs to be determined. For this calculation it is assumed that the coefficient of friction (µ) equals 0.58, we then can determine the force due to friction (Fµ ) and the total force (Ft ) as follows: Fµ = µw = 0.58 × 1000 = 580 lbs

(10.62)

Ft = Fµ + Fa = 580 + 487 = 1067 lbs

(10.63)

Ft 1067 = = 606 psi Ab 1.76 Ft 1067 = = 1089 psi Pr = Ar 0.98

(10.64)

Pb =

One should note that for a single-rod cylinder, the rod-end pressure is always greater than the blindend, but only with double-rod cylinders having equal rod diameters will the pressure be the same at both ends. Finally, the flow rate required at the blind-end (Qb ) and rod-end (Qr ) may be calculated as follows: 37.5 × 1.76 × 60 Vmax × Ab × 60 = = 17.1 gpm 231 231 Vmax × Ar × 60 37.5 × 0.98 × 60 = = 9.6 gpm Qr = 231 231

Qb =

(10.65)

10.2.5 Hydraulic Fluid Considerations 10.2.5.1 Foaming Most hydraulic fluids have an antifoaming agent as an additive. These additives have caused discussions among hydraulic system designers and users. Most of the additives used to control the foaming tendencies of hydraulic fluids accomplish this task by increasing the surface tension of the fluid. When the surface tension increases the size of air or vapor bubbles, which will coexist in the fluid, become smaller and therefore are less likely to rise to the surface and cause a foaming situation. However, when the air is allowed to remain in the fluid, the compressibility of the fluid increases, or stated in another way the bulk modulus of the fluid decreases. The suspension of air or vapor in the circulating fluid of a hydraulic system is a fault of the system. That is, a well-designed system will not permit air or vapor to become entrained in the fluid. Some expert designers of hydraulic systems have said that they would rather not have an antifoaming agent present. Without the addition of the antifoaming agent a system, which is poorly designed, will be readily apparent and can be fixed. Details on foaming, air entrainment, and air release are provided in Chapter 2. 10.2.5.2 Bulk Modulus The bulk modulus of a fluid is a term used to describe the compressibility of the fluid. In fact, the bulk modulus is inversely proportional to the compressibility. The purpose of a hydraulic system is to raise the potential energy of the system by increasing the pressure of the fluid. This potential energy can then be converted into kinetic energy, which will due useful work. However, a fluid with a low bulk modulus will be very compressible and the energy necessary to raise the pressure must also be sufficient to compress the fluid. Most hydraulic fluids have a very high bulk modulus in the pristine condition. However, when air is present, the effective bulk modulus will be low and the system fluid will need to absorb the heat generated when the compression takes place. Calculation procedures for bulk modulus and fluid compressibility are described in more detail in Chapter 2.

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Handbook of Lubrication and Tribology

S

DS DL

V1

A1

To tank

VP

From pump

FIGURE 10.59

Illustration of a cylinder meter-in circuit.

10.2.5.2.1 Fluid Compressibility and Cylinder Lunge Fluid compressibility has a great effect on cylinder performance. Especially when the fluid type is changed, such as changing from a mineral oil to a water-based or synthetic fluid. Hydraulic cylinders are especially sensitive to changes in bulk modulus. In critical operations it is often necessary to extend the cylinder smoothly and at a very constant velocity. If the load changes, the compressibility of the hydraulic fluid will have a negative influence on the constant velocity. Also, any change in the volume ( V ) of the fluid under compression will translate into a change in cylinder stroke ( S) defined as “lunge.” The following expressions can be used to calculate “lunge” ( S) and the resultant velocity change ( v): S =

[Vp + (A × S )] × L A2 β S × 60 v = τ

(10.66) (10.67)

where S is lunge (in.), Vp is volume in pipe (in.3 ), A is effective piston area (in.2 ), S is stroke (in.), L is load change (lbs.), β is bulk modulus of fluid, v is velocity change (in./min), τ is load change time (sec). We will now apply these equations to the meter-in (Figure 10.59) [25] and the meter-out (Figure 10.60) [25] circuits under the following conditions: A1 = 4.9 in.2 (Blind-end area) A2 = 2.5 in.2 (Rod-end area) S = 24 in. (Stroke) L1 = 3000 lbs (Full load) L2 = 1000 lbs (Reduced load) L = 2000 lbs (Load change, L1 − L2 ) τ = 1 sec (Load change time) Vp = 36 in.3 (Oil line volume) β = 200,000 lb/in.2 (Bulk modulus of oil)

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Basic Hydraulic Pump and Circuit Design

10-51

DS

S

V2 DL

A2 From pump VP

To tank

FIGURE 10.60

Illustration of a cylinder meter-out circuit.

For the meter-in mode (Figure 10.59) using Equations (10.66) and (10.67) we calculate: [36 + (4.9 × 24)] × 2000 = 0.064

(4.9)2 (200,000) 0.064 × 60 = 3.8 in./ min v = 1

S =

(10.68) (10.69)

For the meter-out mode (Figure 10.60) using Equations (10.66) and (10.67) we calculate: S =

[36 + (2.5 × 24)] × 2000 = 0.154

(2.5)2 × 200,000 0.154 × 60 = 9.2 in./ min v = 1

(10.70) (10.71)

As you can see from the examples above, the degree of “lunge” is directly proportional to the load change and inversely proportional to the bulk modulus of the fluid. In addition, cylinder lunge is greater in the meter-out mode than in the meter-in mode. This is due to pressure intensification in the rod-end of the cylinder as discussed earlier in this chapter (Section 2.1.5).

References [1] Frankenfield, T.C., Using industrial hydraulics, Chapter 1, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [2] Blackburn, J.F., Reethof, G., and Shearer, J.L., Fluid Power Control, The M.I.T. Press, Cambridge, MA, 1960. [3] Fitch, E.C. and Hong, I.T., Hydraulic Component Design and Selection, BarDyne Inc., Stillwater, OK, 1997.

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[4] Frankenfield, T.C., Using industrial hydraulics, Chapter 6, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [5] Sullivan, J.A., Fluid Power Theory and Applications, 2nd ed., Reston Publishing Company, Reston, Virginia, 1982. [6] Mannesmann Rexroth, Pump and Controls — Open Loop, The Rexroth Training Center, Bethlehem Vocational Technical School, Bethlehem, PA, June 1995. [7] Fitch, E.C., Fluid Contamination Control, FES Inc., Stillwater, OK, 1988. [8] Tessmann, R.K. and Howsden, J.M., Environmental influence upon wiper seal performance, BFPR Journal, P75-5, FPRC/OSU, Stillwater, OK, October 1975. [9] Tessmann, R.K. and Howsden, J.M., Service life of wiper seals, BFPR Journal, P76-31, FPRC/OSU, Stillwater, OK, October 1976. [10] Fitch, E.C. and Tessmann, R.K., Modeling the performance of filter assemblies, BFPR Journal, P73-CC-12, FPRC/OSU, Stillwater, OK, October 1973. [11] Fitch, E.C. and Tessmann, R.K., The filter selection graph — a basic contamination control tool, BFPR Journal, P74-55, FPRC/OSU, Stillwater, OK, October 1974. [12] Wolf, M.L., Contaminant Particle Effects on Pumps as a Function of Size, Type and Concentration, M.S. Thesis, FPRC/OSU, Stillwater, OK, 1965. [13] Tessmann, R.K., Contaminant wear in hydraulic and lubricating systems, BFPR Journal, P75-4, FPRC/OSU, Stillwater, OK, October 1975. [14] Bensch, L.E., Verification of the pump contaminant wear theory, BFPR Journal, 11, FPRC/OSU, Stillwater, OK, October 1977. [15] Hydraulic Hints and Troubleshooting Guide, No. 694, Vickers, Incorporated, Troy MI, August, 1996. [16] Fitch, E.C., Fluid Power Engineering, FES Inc., Stillwater, OK, 1982. [17] Mackay, R.C., Pump suction conditions, Pumps and Systems Magazine, 20, May 1993. [18] Paul-Munroe, Lightning Reference Handbook, 8th ed., Rucker, Inc., 1994. [19] Frankenfield, T.C., Using industrial hydraulics, Chapter 4, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [20] Frankenfield, T.C., Using industrial hydraulics, Chapter 5, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [21] Norvelle, F.D., Fluid Power Technology, West Publishing Company, New York, NY, 1995. [22] Frankenfield, T.C., Using industrial hydraulics, Chapter 10, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [23] Frankenfield, T.C., Using industrial hydraulics, Chapter 9, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [24] Hong, I.T. and Tessmann, R.K., What time do you have? Proceedings of the National Conference on Fluid Power, Vol. II, National Fluid Power Association, pp. 23–25, 1996. [25] Frankenfield, T.C., Using industrial hydraulics, Chapter 3, Hydraulics and Pneumatics Magazine, Cleveland, OH, 1990. [26] Wilson, W.E., Rotary-Pump Theory, Transactions of the A.S.M.E., 68, 371–384, May 1946. [27] Wilson, W.E., Clearance Design in Positive-Displacement Pumps, Machine Design, 127–130, February 1953. [28] Handbook of Chemistry and Physics, The Chemical Rubber Co., 49th ed., p. D109.

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11 Hydraulic Fluids 11.1 Functions of Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . 11.2 Types of Hydraulic Fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

11-1 11-2

Fluids in General • Fire-Resistant Fluids • Environmentally Compatible Fluids • Special Fluids

11.3 Properties of Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . .

11-7

Viscosity • Density • Compression Modulus • Load-Bearing Capacity and Antiwear Properties • Setting Point/Pour-Point • Gas Solubility • Air Release Capacity and Foaming Tendency • Aging Behavior • Material Behavior • Detergent and Dispersant • Thermal Capacity and Thermal Conduction • Flammability • Biodegradability and Toxicity

H. Murrenhoff and O.-C. Göhler Institute of Fluidpower Drives and Controls (IFAS) RWTH Aachen University

T. Meindorf Argo-Hytos GmbH

11.4 Fortifying Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . . . .

11-18

Oxidation Inhibitors • Metal Deactivators • Wear Protection • Friction-Reducing Agents, Friction-Modifiers • Viscosity Index Improvers • Setting Point Depressants • Antifoaming Agents • Detergents and Dispersants • Corrosion Inhibitors

11.5 Impurities in Hydraulic Fluids . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

11-22 11-23

Hydraulic fluid is a complex component which is part of the hydraulic system. Apart from its primary function of transferring energy, the medium also has to perform a number of other functions and is predominantly used as a lubricant. The characteristic profile required for this can no longer be fulfilled by a base fluid based on mineral oil, as it used to be in the early days of oil hydraulics. The technical performance capability is now supplemented or extended by special chemical additives. Other base fluids can be used to meet additional requirements where a fluid is required to be fire resistant, environmentally compatible, or capable of withstanding extreme loads, for example.

11.1 Functions of Hydraulic Fluids The primary function of a hydraulic fluid is to transfer energy from a pressure generator to the consumer. A volume connection is established for this, which is analogous to the form closure or frictional connection in a mechanical transmission concept. After meeting this functional requirement, the hydraulic medium has to perform other functions. It acts as a lubricant, thereby reducing the friction and wear of parts moving against one another in 11-1

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Handbook of Lubrication and Tribology

Power transmission Pressure transmission

Volume connection Functions of hydraulic fluids

Heat removal

Corrosion protection

FIGURE 11.1

Wear reduction

Elastomer compatibility

Functions performed by hydraulic fluids.

bearing arrangements, for example. The components of the hydraulic system must be protected against corrosion and other chemical reactions. The losses attributable to friction and throttling are discharged by the hydraulic medium in the form of thermal energy. Figure 11.1 shows an overview of the functions performed by the fluid within the hydraulic system. These functional properties of the fluid must be assured over a wide temperature range. The characteristics of the medium itself must not be impaired by impurities, such as rubbed-off particles or water. After all, there are both economic and ecological arguments in favor of a long lifetime for the medium with its properties remaining unchanged. Safety aspects give rise to additional requirements in situations where the fluid must not be inflammable or must have a reduced evaporation tendency for reasons related to working safety. Increasing importance has been attached to ecological aspects for some years now. As far as mobile applications in particular are concerned, there are growing demands for fluids which, unlike conventional media, decompose quickly and completely in the event of leakage and are less harmful in toxicological terms. The requirements to be met by hydraulic fluids are summarized in Figure 11.2.

11.2 Types of Hydraulic Fluid Hydraulic media generally comprise a base fluid which is doped to produce a ready-to-use formulation with the required properties by adding other substances, which are referred to as additives. The proportion of these additional substances ranges from just a few percent in lightly doped mineral oils right through to 50% in a few fire resistant fluids. The nature of the base fluid, which in itself may be a mixture of heterogeneous molecules of the same type, for example ester molecules with a varying degree of saturation of the fatty acid, essentially determines the application for which the hydraulic fluid is used.

11.2.1 Fluids in General Accounting for approximately 80 to 85% of hydraulic fluids, those based on mineral oil constitute the most significant group of hydraulic media in terms of quantity and cost effectiveness. They are capable of meeting a broad spectrum of requirements and therefore find universal application in stationary and mobile systems. Refined by many years of experience with these media — accompanied by as many years spent in their development — the properties of these media are clearly defined and can be reproduced at any time.

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-3

High EP characteristics

Low compressibility

High aging stability Demands on hydraulic fluids

High heat capacity and conductivity

Non corrosive

FIGURE 11.2

High biodegradability

Low flammability

Requirements to be met by hydraulic fluids.

Table 11.1 shows the mineral oil sales figures for Germany for 1997–2003, which have been published by the Association of the German Oil Industry, Mineralöl Wirtschafts Verband (MWV). The base fluids are produced by distilling and refining crude oil and usually comprise a mixture of C20 to C35 molecules, consisting of normal paraffin, isoparaffin, naphthene, and isolated aromatic compounds. The essential properties of the fluid, such as viscosity, flash point, aging stability, viscosity/temperature behavior, and response to low temperatures are determined by varying the stages in the process [1]. Special properties not demonstrated by the base oil, or not to an adequate extent, are achieved by means of additives. The types of doped additives are distinguished according to the classification of the fluid. This means, for example, that hydraulic fluids which have no active ingredients and are hardly used nowadays in practical terms belong to Group H. Group HL contains hydraulic fluids with active ingredients which improve aging stability and corrosion protection and Group HLP contains hydraulic fluids with additional active ingredients which reduce wear and increase the capability of withstanding load. Hydraulic fluids containing additional active ingredients which increase the viscosity/temperature index (VI) to a value in excess of 140 are found in Group HVLP. This elevation can also be achieved in the base oil by means of a special refining process known as hydro-cracking. This classification is specified by DIN 51 524 “Pressure fluids — Hydraulic oils; Minimum requirements,” which also stipulates the minimum requirements to be met with respect to viscosity characteristics and aging stability, for example, as well as corrosion protection and antiwear properties. Additives with detergent and dispersant effects are added to the fluids for special applications. These types of fluid are identified by the letter D in the nomenclature. The additives in an HLP-D fluid, for instance, render it capable of binding up to 5% of water in the form of an emulsion. The water is enclosed in the form of tiny droplets in such a way as to avoid obstructing the formation of a lubricating film, with the result that no corrosion occurs on the metallic surfaces and there is no intensified aging of the medium. These fluids are given precedence in maritime applications, where there is a particularly high risk of the medium being contaminated by water. Table 11.2 shows an overview of the currently applicable mineral oil standards.

11.2.2 Fire-Resistant Fluids Fluids which are classified as being fire-resistant have been developed for use in the mining, aircraft construction, die-casting, and rolling mill industries. These fluids have a much higher inflammation temperature than mineral oils and their use is mandatory for certain applications. This means, for example, that the fire prevention regulations for underground mining forbid the use of mineral oil. A distinction

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Handbook of Lubrication and Tribology

TABLE 11.1

German Lubricant Sales Figures Classified According to Grades, 1997–2003

Groups of substances

1997

1998

1999

2000

2001

2002

Engine oil Compressor oil Turbine oil

411.5 9.3 4.5

381.6 9.9 5.2

384.9 11.1 5.2

369.9 10.9 3.6

343.7 10.5 2.7

348.7 12.0 2.6

343.9 11.6 2.3

Transmission fluid Automotive Industrial Hydraulic fluid

66.9 24.1 153.9

64.0 26.7 158.2

66.9 28.9 150.3

66.0 28.0 154.6

68.1 24.8 146.0

70.3 22.1 143.9

68.0 21.1 135.5

46.2 31.4 3.0 9.9

48.2 30.7

45.4 29.8

45.6 32.8

47.2 29.7

50.4 28.1

10.3

9.5

9.2

7.2

7.6

49.6 27.9 2.0 7.9

42.4 9.0 99.3 18.8 50.5 59.6

38.6 9.4 80.3 19.2 34.4 59.9

49.3 10.0 85.4 15.1 35.7 54.9

44.4 10.9 83.3 10.9 30.8 53.4

38.6 16.8 84.8 12.6 30.3 55.0

34.8 17.9 85.8 13.7 32.2 51.7

45.4 11.2 108.9 9.1 31.4 48.0

32.4 45.9 49.4

34.4 44.4 63.5

35.7 45.2 83.0

30.8 45.9 72.0

30.3 43.7 55.2

32.2 44.2 79.0

35.9 33.8 73.1

1168.0 75.0

1146.8 73.0

1159.9 51.8

1122.3 77.1

1057.7 96.1

1076.6 110.2

1066.8 101.6

Metal working oil Straight Water-miscible Hardening oil Anticorrosive agents Paraffin oil Medical Technical Other process oils Electrical insulating oil Machine oil Other industrial oils/fluids not used for lubrication Lubricating grease Extracts from lubricating oil refinery Base oil Total Including quantities of recycled waste oil (already included in the respective groups)

2003

Source: Federal Office for Trade and Industry.

TABLE 11.2

Mineral Oil Standards

DIN 51 524

ISO 6743-4

H

HH

Without special additives (base oil)

HL

HL

HLP

HM

HVLP

HV

HLPD

(-)

With active ingredients to increase corrosion protection and aging stability. DIN 51 524, Part 1 As for HL but with further additives to reduce fretting wear with mixed friction. DIN 51 524, Part 2 As for HLP but with further additives for improve the viscosity/temperature behavior. DIN 51 524, Part 3 As for HLP but with further additives to dissolve deposits (detergent) and waterbearing to a certain extent (emulsifying/ dispersing)

Composition

Fields of application Systems with no special requirements (infrequent) Systems with moderate pressures, but high temperatures. Satisfactory separating capacity Systems with high pressures and temperatures. High-quality, widely used hydraulic fluid, particularly HLP 46 Wider temperature range than HLP with low initial values as a consequence of shallow viscosity characteristic Systems where water may ingress to fluid fill (condensation, cooling lubricants for machine tools, mobile systems)

is made between aqueous and nonaqueous hydraulic fluids. VDMA Guidelines 24 317 and 24 320 classify these fluids as follows (the designations are in accordance with CETOP RP 77 H, ISO 6743, Part 4, and DIN 51 502). The HFA group contains oil in water emulsions or highly aqueous solutions, the HFB group contains water-in-oil emulsions, the HFC group contains aqueous solutions, and the HFD group contains nonaqueous fluids.

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Hydraulic Fluids

11-5

The concentrate component of HFA fluids usually accounts for between 1% and a maximum of 5%, the rest is water. The permitted 20% limit is not exploited, for reasons related to costs. The concentrate component may be either an emulsifiable substance or a medium that dissolves in water. The concentrate contains additives which improve anticorrosion and antiwear properties or increase lubricity. It also contains biocides which prevent the formation of bacteria, fungi, and yeast in the medium. This type of fluid is primarily used to support mine workings hydraulically. Other applications include hot-working and manufacturing facilities in the automotive industry. HFB fluids comprise 40% water and 60% mineral oil. Unlike the situation in British collieries, this type of fluid is not used in Germany as it does not pass one of the fire tests prescribed by the German mining authorities. HFC fluids comprise 35 to 55% water, and viscosity-increasing substances and other additives make up the remaining 45 to 65%. In most cases, polyakylene glycols are used as thickening agents in order to increase the viscosity to a level that is similar to that of mineral oil. Anticorrosives and antiwear substances are used as the additives. This type of fluid is primarily used in high-performance underground mining systems and in hot-working plants. HFD fluids are nonaqueous, synthetic media. Phosphate esters and chlorinated hydrocarbons, the chloroaromatic compounds, are most frequently used. Diesters and silicone oils, polyphenylene ethers, polyglycols, silicate esters, and fluorocarbons are also used for special applications. These fluids are identified more precisely by an extra letter added to the “HFD.” Because of their environmental impact, fluids containing chlorinated hydrocarbons (PCBs) are used only under exceptional circumstances. The applications for this type of fluid include hot-working plants, such as die-casting machines, as well as mining. An overview of the currently applicable standards for fire resistant hydraulic fluids is given in Table 11.3.

11.2.3 Environmentally Compatible Fluids Efforts to substitute mineral oil products with renewable, plant-based raw materials were already being made — predominantly in Finland — during the first oil crisis at the beginning of the 1970s. However, the development of ecologically friendly hydraulic fluids is primarily oriented to a high level of environmental compatibility, and these fluids are being used to an increasing extent in mobile systems as well as stationary installations. In Germany, VDMA Guideline 24 568 specifies the minimum requirements for four classes of fluid: HETG (hydraulic oil environmental native triglycerid), HEES (hydraulic oil environmental synthetic ester), HEPG (hydraulic oil environmental polyglycol), and HEPR (hydraulic oil environmental polyalphaolefins and related products). The oils obtained from rapeseed and cognate plants possess very good technical properties, have a high VI, and, as such, constitute an ideal basis for HETG “native” hydraulic fluids. Like all vegetable oils these contain an ester molecule, which comprises a trivalent glycerine as the alcohol component and three

TABLE 11.3

Standards for Fire Resistant Fluids

ISO 6743/CETOP Lux. Ber./VDMA HFA HFB HFC

Composition Oil-in-water emulsion or synth. aqueous solution containing max. 20% concentrate Water-in-oil emulsion containing max. 60% oil Aqueous polymer solution containing 35 to 55% water

HFDU

Carboxylic ester (nonaqueous, synthetic)

HFDR

Phosphate ester (nonaqueous, synthetic)

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Fields of application Mining, hydraulic presses, temperature range: 5 to 55◦ C Mining, temperature range: 5 to 60◦ C Mining, foundries, moderate pressure, environmental protection, temperature range: −20 to 60◦ C Temperature range: −35 to 100◦ C, more widely used than HFDR Motor vehicles, aerospace industry, temperature range: −20 to 150◦ C

11-6

Handbook of Lubrication and Tribology

fatty acids, such as oleic acid, linoleic acid, or erucic acid. The typical product for a plant is determined by the proportionate quantities of the fatty acids. The term “esterification” is used to refer to the reversible process which takes place during the reaction of alcohols and acids, as given in a simplified manner in the following equation: Alcohol + Acid
Ester + H2 O

(11.1)

Esters are polar in character. The resulting adhesion to metal surfaces is accompanied by excellent antiwear properties and the surface wetting process also provides satisfactory protection against corrosion [1–5]. The ester compounds and the characteristic double carbon bonds of the fatty acids are changed relatively easily by means of oxidation (adding oxygen), hydration (adding hydrogen), or hydrolysis (breaking down into alcohol and free fatty acids). As a result, the native oils demonstrate low aging stability and give rise to aggressive cleavage and reaction products under the very conditions, with the associated high temperatures, which are usually encountered in service. On the other hand, these properties also enable rapid and virtually complete biodegradation of any fluid leaking into the ground, bodies of water, or sewage plants. This means that native oils must be fortified by suitable substances to prevent the fluid aging while in service. Of course, the additives used for this must not hinder the natural biological decomposition and must be nontoxic in themselves [1,4]. Synthesis can be used to produce esters which are similar to vegetable oils in terms of chemical structure and ecological characteristics but demonstrate much better aging stability. In this respect, there are diverse possible combinations of alcohols and acids which offer a means of selectively formulating products which meet the requirements for HEES hydraulic media [4]. Natural or petrochemical products are chosen as the starting materials and these are combined in different ways in the subsequent stages of the process. However, molecular stabilization is generally accompanied by a reduction in the rapid biodegradation capacity. Additives must be mixed into these fluids as well, and the additives are also subject to the ecological restrictions. A distinction is made between two main groups of synthetic esters which constitute suitable base fluids for hydraulic media. Natural fatty acids, and oleic acid (C 18 : 1) in particular, are used for unsaturated esters, such as TMP (trimethylolpropane) esters. Although their double bonds guarantee good lowtemperature properties, they detrimentally affect aging stability. The chemical structure is very similar to that of the native oils. High-quality, completely saturated esters without double bonds have predominantly petrochemical origins. Their structure differs from that of the native oils to a considerable extent. Rapidly degradable, nontoxic base fluids can be produced with properties that even surpass those of mineral oils. However, the manufacturing process is very complex and very expensive. Polyglycols have been used as high-performance synthetic lubricants for several decades now. Of this group, the polyalkylene glycols present significantly less environmental hazard potential than mineral oils and are therefore used as base fluids for the HEPG class. In the petrochemical synthesis process, the viscosity can be influenced by varying the relative molecular mass, which means that there are different viscosity classes available. The viscosity’s temperature dependence is less pronounced for ester fluids than it is for mineral oils with a VI of 185 to 215 without any fortification. The tribological properties are equivalent to or better than those of the mineral oils. Unlike oils, polyglycols are water soluble. This plays an important role in their biodegradation, which can take place in the solution rather than in the boundary surface, as is the case for oils. Rain water quickly carries any leakages into the deeper, lifeless earth layers where there is less oxygen so that the polyglycols can get into the groundwater without being decomposed beforehand. This is why they are no longer regarded as being environmentally compatible in Austria. Fluids which are predominantly synthesized from polyalphaolefins (PAOs) and cognate hydrocarbons are classified as belonging to a relatively new group with the designation HEPR. While PAOs are frequently used in the transmission fluid sector, the HEPRs are not yet playing a significant role in hydraulic systems.

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-7

Nor has any long-term experience been gathered with this type of oil. One of the unclarified aspects is their compatibility with the materials usually used in hydraulic systems. One factor that must be regarded in a critical manner is that the HEPRs may include multifarious formulations with virtually any combinations of PAOs and hydrocarbons, which makes it difficult to agree on a choice of materials. When environmentally compatible fluids are used, the fact is that they may react differently with plastics and certain metals than mineral oil. The materials used for seals and hoses, as well as sliding bearings and filter elements, for example, must therefore be matched to the fluid. If a polyglycol is used, it is usually necessary to remove the coating on the inside of the container as polyglycols dissolve paint and lacquer. The relatively high prices of synthetic esters, particularly those of saturated synthetic esters, are still preventing the HE media from being used in any significant quantities. A study of all lubricants and hydraulic media showed that bio fluids accounted for just 1.8% of the total lubricant market in Germany in 2003. Only around 36% of these belong to the group of biogenic fluids, that is, those based at least 50% on renewable raw materials. Although the remaining proportion of bio fluids are rapidly biodegradable and ecotoxicologically harmless, these are still based on beef suet and petrochemical products to the greatest extent. If the focus is on the hydraulic fluids alone, around 6.9% of the lubricants and hydraulic fluids can be described as bio oils. Biogenic oils account for 2.4% of the market as a whole. The bio oils (>90%) are almost exclusively found in the mobile hydraulic sector (approximately 60,000 Tonnes per annum). Only a very small proportion are used in stationary installations. In 2003, 17% of mobile hydraulic systems were using bio oils, while biogenic fluids accounted for 6%. Official stipulations and requirements, financial benefits resulting from the use of less water-polluting substances, and the users’ increasing environmental awareness are leading to noticeable expansion in the market.

11.2.4 Special Fluids Apart from the aforementioned media which can be used for more or less any purpose within their respective field of applications, there are many fluids that have been developed for specific, primarily mobile applications. The high temperatures prevailing in car and truck braking systems require highly stable base fluids, usually polyalkylene glycol. These fluids must be capable of absorbing a lot of water as undissolved water can boil prematurely in the brake line, drastically reducing the braking effect. The hydraulic systems, steering, transmission and wet brakes of a farming tractor are all supplied from a common fluid reservoir. The universal tractor oils used for this, with the designations UTTO and STOU, therefore unite the properties of a hydraulic medium with those of a transmission lubricant. The automatic transmission fluids (ATFs) used in torque converters and automatic transmissions fulfill similar requirements. They have been tuned to such an extent as to achieve constant frictional behavior as stick-slip effects are not wanted in multi-plate clutches. Flow improvers are added to the ATFs to ensure that they maintain their full functional capability during the cold winter months. Preference is given to the use of clarified water as a hydraulic medium where hygiene is important in the food industry. In these cases, the components have to perform virtually all of the tribological functions. Problems are also encountered as a result of the high cavitation tendency and the internal leakage caused by the low viscosity. The performance level has been improved here by the most recent developments in the material and manufacturing technology sectors.

11.3 Properties of Hydraulic Fluids This section describes the properties and parameters which are characteristic of a hydraulic fluid. The approximate values for various groups of fluids are given in the form of Table 11.5 at the end of the chapter.

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Handbook of Lubrication and Tribology

11.3.1 Viscosity The dynamic viscosity is the most important parameter in describing the load-bearing capacity of a hydraulic fluid. Conventional viscometers do not measure the dynamic viscosity directly, but rather its relationship to the density of the medium. This kinematic viscosity is defined as ν=

η ρ

(11.2)

The fluids are divided into certain classes according to their respective kinematic viscosity values. This classification is based on a mean viscosity at 40◦ C in accordance with DIN 51 519, which is the ISO viscosity classification for industrial liquid lubricants. This standard lays down a permanently specified series of numbers for the viscosity values. This series of numbers is as follows: 10, 15, 22, 32, 46, 68, 100 … mm2 /sec. The viscosity may deviate from this value by a certain amount, for example, by ±10 mm2 /sec for viscosity class ISO VG 100. As far as the unit of notation is concerned, the old unit centiStokes (cSt) is still used, as well as mm2 /sec. A fluid of viscosity group (VG) 46, for instance, is identified by adding the number 46 to its designation (e.g., HLP 46). The viscosity is highly dependent on the temperature for the majority of hydraulic fluids. The viscosity is plotted according to Ubbelohde in Figure 11.3. The double logarithm of the viscosity is plotted in mm2 /sec + 0.8 along the y-axis and the logarithm of the temperature is plotted along the x-axis. This type of graph gives rise to straight lines for mineral oil and HFC fluids over a relatively wide temperature range. The shallower the characteristic, the more favorable the cold-start behavior and operating viscosity. This applies to a particular extent in mobile applications which have to cover a broad temperature range between startup at temperatures below freezing and full-load operation during the hot summer months. The characteristic for HFD media rises very progressively in the lower temperature range, where its viscosity is more highly dependent on temperature than that of the other fluids. Shallow in other respects, the characteristic for HE media rises noticeably at low temperatures due to partial crystallization of the molecules. The ascending gradient of the fluid characteristic in the Ubbelohde diagram serves as a judgment scale for the viscosity/temperature (VT) behavior, which is referred to as the VI. A high VI indicates a low dependence on temperature. The VI can be increased by means of special additives, the VI improvers. 10000 [mm2/sec]

HLP HFC/HEES HFD HETG

Kinematic viscosity h

1000

100

10

3 253

FIGURE 11.3

273

313 293 Temperature q

Kinematic viscosity as a function of the temperature.

© 2006 by Taylor & Francis Group, LLC

333

353 [K]

373

Hydraulic Fluids

11-9 10000 [mPasec] 5000

HFD HLP

Dynamic viscosity h

2000 1000 HETG/ HEES

500 200 100

HFC

50 20 10 0

1000

2000

3000 [bar]

4000

Pressure p

FIGURE 11.4 Dynamic viscosity of various hydraulic media as a function of pressure with the same initial viscosity.

HLP has a viscosity index of around 100; HV has a VI of up to 400; the VI for HFC is around 150 and the VI for HFD is even less than zero. Hydraulic media based on vegetable oil have a very high inherent VI of around 200. The viscosity/pressure behavior of a medium is essentially responsible for the load-bearing capability of a liquid lubricating film. A common property shared by all media is that the dynamic viscosity increases when they are subjected to pressure. Figure 11.4 shows this behavior as demonstrated by different hydraulic media. The following equation applies to the pressure dependence of the dynamic viscosity at a constant temperature: η = η0 · eb·p

(11.3)

where η0 is dynamic viscosity at atmospheric pressure, b ≈ 1.7×10−3 bar−1 for mineral oil, b ≈ 3.5×10−4 bar−1 for an HFC medium, b ≈ 2.2 × 10−3 bar−1 for an HFD medium, and b ≈ 1.1 × 10−3 bar−1 for an HETG or HEES medium. Assuming an increase in pressure from 0 to 2000 bar, the viscosity of the HFC fluid increases by factor of 2, that of mineral oil by factor of 30, and that of the HFD fluid by factor of 80. The increase in viscosity as the pressure rises exerts a positive influence under high bearing loads as the lubricating film undergoes a self-reinforcing effect. The behavior shown in Figure 11.4 is one of the reasons why rolling bearings have a relative short service life when water-based fluids (HFA, HFC, clarified water) are used. The viscosity cannot be increased to any significant extent by means of additives.

11.3.2 Density The losses in pipelines and the flow channels of components are directly proportional to the density of the hydraulic medium. The density is also required to calculate the dynamic viscosity from the kinematic viscosity, as described above. The density is temperature-dependent as the volume of a fluid expands when the temperature increases. The coefficient of expansion (γ ) is defined as γ = where ϑ is temperature.

© 2006 by Taylor & Francis Group, LLC

1 ∂V V ∂ϑ

11-10

Handbook of Lubrication and Tribology

This gives rise to an initial volume V0 with an increase in temperature by ϑ for a change in volume V : V = V0 γ ϑ

(11.4)

The density of the fluid decreases in proportion to the increase in volume. m V0

ρ0 =

m V0 + V

ρ= give rise to the following equation for the density: ρ=

ρ0 1 + γ ϑ

(11.5)

This equation relates to 15◦ C for the normal range of applications. Values for ρ15◦ C and γ can be found in the table at the end of this chapter. Figure 11.5 shows the density of various standard hydraulic fluids as a function of the temperature. The coefficient of expansion for mineral oil amounts to 7 × 10−4 K−1 . This means that the volume expands by 0.7% when the temperature increases by 10◦ C.

11.3.3 Compression Modulus The density of a hydraulic fluid is determined by the pressure. This compressibility is very important for the dynamic performance of hydraulic systems. The compressibility coefficient β is defined as 
1 ∂V 1 β=− = (11.6) V0 ∂p ϑ EFl where EFl is the compression modulus.

1200 HFD

[kg/m3] 1100

Density r

HFC HFA

1000

HEES/HETG 900 HLP

800

FIGURE 11.5

0

20

40

60 80 Temperature q

Density of hydraulic fluids as a function of the temperature.

© 2006 by Taylor & Francis Group, LLC

100

120 [°C]

Hydraulic Fluids

11-11

The density of the fluid increases according to the reduction in volume when the pressure rises. ρ=

m V

V = −V0 β p gives rise to ρ=

ρ0 1 − β p

(11.7)

where ρ0 is the density at atmospheric pressure. This ratio applies to the usual pressure and temperature conditions in hydraulic systems. A mean constant compressibility coefficient β can be used for calculation here, even though β actually decreases as the pressure rises and increases with the temperature. Analogous to the modulus of elasticity for solids, the reciprocal value of β is referred to as the compression modulus EFl : EFl =

1 dp = −VFl,0 β dVFl

(11.8)

Accordingly, the compression modulus increases with the pressure and decreases with the temperature. Figure 11.6 shows the true adiabatic compression modulus of a mineral oil as a function of the pressure with the temperature plotted as a parameter. The term “true adiabatic compression modulus” is used to refer to the ascending gradient of the volume/pressure curve at the respective pressure values. Undissolved air components of 5 to 10% by volume are often found [9], particularly in mobile systems with short circulation times and unfavorable tank designs. As shown by the calculation below, these gas bubbles exert a very strong influence on the compressibility of the fluid and therefore on the stiffness of the system under load and its dynamic performance. The volume of a fluid/air mixture V0 = VFl,0 + VL,0 is subjected to a change in pressure dp, where VFl indicates the volume of the fluid and VL is the volume of dissolved air. The equivalent compression modulus EG of the mixture is given by EG = −

VFl + VL (dVFl /dp) + (dVL /dp)

(11.9)

True adiabatic compression modulus EFl [104 bar]

4 °C 10 30

3 50 70 90 110

2

1

0

700 Pressure p

[bar]

FIGURE 11.6 True adiabatic compression modulus of an HLP 46 mineral oil.

© 2006 by Taylor & Francis Group, LLC

1400

11-12

Handbook of Lubrication and Tribology

The compression modulus of the fluid up to a pressure of 700 bar is approximately linearly dependent on the pressure (also refer to Table 11.3) EFl = E0 + m · p

(11.10)

Equation 11.8 and p0 = 0 bar give VFl = VFl,0 · e−(1/EFl )(p−p0 ) VFl,0 dVFl =− dp E0


1+

m·p E0

(m+1)/m

(11.11)

The air is assumed to be an ideal gas, so that the following is given for polytropic changes in state: n p · VLn = const = p0 · VL,0

VL,0 dVL =− dp n · p0




p0 p

(n+1)/n

(11.12)

The equivalent compression modulus of a fluid/air mixture is therefore given by EG =

VFl,0 (1 + (m · p)/E0 )−1/m + VL,0 (p0 /p)1/n (VFl,0 /E0 )(1 + (m · p)/E0 )−(m+1)/m + (VL,0 /(n · p0 ))(p0 /p)(n+1)/n

(11.13)

If α is then defined as the air content of the mixture in its initial state: α=

VL,0 V0

the equivalent compression modulus can be calculated from: EG =

(1 − α)(1 + (m · p)/E0 )−1/m + α(p0 /p)1/n (1/E0 )(1 − α)(1 + (m · p)/E0 )−(m+1)/m + (α/(n · p0 ))(p0 /p)(n+1)/n

(11.14)

n = 1 for slow changes in state with isothermal characteristics. Figure 11.7 shows the equivalent compression modulus characteristic for a mixture of mineral oil and air compared with air-free fluid for α = 0.1, 1, and 10%, E0 = 15,000 bar, and m = 10. This calculation does not allow for the air-dissolving capacity as a function of the pressure. NFPA Standard T2.13.7R1-1997 provides a procedure for determining the compressibility of a fluid [10]. If the gas content is low, the mixture already reaches the compression modulus of the fluid at around 50 bar. A higher gas content leads to an increase in compressibility throughout the entire normal operating pressure range. Another effect can also occur here, and that is, that the change in state — for example, when the fluid is being pumped — takes place very quickly and is therefore approximately adiabatic. As a result, the gas bubbles suddenly heat up to a high temperature. In extreme cases, the temperature reaches the inflammation temperature of the fluid and the fluid is damaged by what is referred to as the “micro-diesel effect.” The gas release and redissolving behavior of the various types of media and other influences are such that the actual characteristic of the compression modulus deviates from the values calculated above. Experiments must therefore be carried out in order to be able to calculate the dynamic performance of the system exactly.

© 2006 by Taylor & Francis Group, LLC

Equivalent compression modulus EG [104 bar]

Hydraulic Fluids

11-13 2 Proportion of air a: 0.1%

1.6 1%

10%

1.2

0.8 Compression modulus for the fluid Isothermal change in state Adiabatic change in state

0.4

0 0

50

100

150

200

250 [bar]

300

Pressure p

FIGURE 11.7 Compression modulus of a mixture of mineral oil and air.

11.3.4 Load-Bearing Capacity and Antiwear Properties A high load-bearing capacity is one of the most important requirements to be met by a hydraulic fluid, and this also implies good antiwear properties. The dynamic viscosity is the most important parameter for the antiwear properties for hydrodynamic lubrication. If the forces acting at low sliding velocities in the mixed friction zone are not sufficient to separate the mating frictional surfaces completely, then the antiwear properties of a fluid are determined by its ability to wet a metal surface and form friction-reducing reaction layers on the mating faces. The wetting capability is also referred to as “oiliness.” The load-bearing capacity of a hydraulic fluid in the mixed friction zone can be improved by means of antiwear additives and substances which reduce the coefficient of friction.

11.3.5 Setting Point/Pour-Point The setting point of a fluid is determined by the temperature at which the medium just ceases to flow under certain testing conditions. By comparison, the pour-point corresponds to the temperature at which the medium just continues to flow. This is around 6 to 8◦ C higher than the setting point. Determination of the pour-point alone is not admissible for an evaluation of the low-temperature characteristics of ester-based hydraulic media. The slow crystallization processes occurring here are such that the time dependence must be taken into consideration as well as just the temperature, and this is determined by means of special test procedures.

11.3.6 Gas Solubility All hydraulic fluids are capable of dissolving a certain proportion of gas. This gas solubility is proportional to the pressure up to around 300 bar. Henry’s law applies: VG = VFl αV

p p0

(11.15)

where VG is volume of gas dissolved at the reference pressure; VFl is volume of fluid; p0 is atmospheric pressure, reference pressure; p is absolute pressure; αV is Bunsen coefficient. The Bunsen coefficient indicates the percentage of gas by volume which is dissolved in a volume unit of the fluid under normal conditions (1.013 mbar, 20◦ C). This value must be determined for every gas or gas mixture. The Bunsen coefficient of air is only slightly dependent on temperature and viscosity. Values for

© 2006 by Taylor & Francis Group, LLC

11-14

Handbook of Lubrication and Tribology

Constant geometric volume

Variable geometric volume

Closed system

q q

x·A q q

Displacement chambers in pumps motors cylinders

Closed tank Closed pipeline

Open system

QIn

QOut Q In

Suction line etc. vIn

vOut

Control edge, screw joint, etc.

FIGURE 11.8

x·A

Displacement chambers in pumps motors cylinders

Buildup of negative pressure in hydraulic systems.

various fluids are given in Table 11.5 at the end of this chapter. The Bunsen coefficient may be determined by ASTM D3827. Under normal circumstances, dissolved air does not exert any influence on the properties of the hydraulic fluid. It may bleed out of the fluid, however, if a low static pressure is applied locally, particularly if the fluid is simultaneously subjected to shearing stresses. The process is referred to as cavitation. The word “cavitation” literally means the formation of cavities. As shown in Figure 11.8, a negative pressure builds up in constant or variable geometric volumes and in open or closed systems. If, for example, a system is shut down with a volume of fluid enclosed in a line, thermal contraction causes the pressure to drop. Where piston port and valve controlled displacement chambers widen in pumps, motors, or cylinders, a partial vacuum negative pressure builds up if the fluid does not continue to flow through to an adequate extent. The gas released as a consequence of this is not redissolved until the pressure increases, when it is dissolved again quickly. Cavitation is most likely to occur in an open system with a constant volume. Such systems include pump suction lines and intake ports, where flow losses are caused by narrow cross-sections, filters, manifolds, and excessive suction height. The consequences are disturbances in the delivery behavior, noise, and an increase in wear due to inadequate lubrication. A low absolute pressure may prevail in flow resistors, for example, throttles, orifices, control edges, even if the system pressure is high. The narrowing cross-section causes the pressure to be converted into a high kinetic energy with simultaneous shearing stresses. These circumstances frequently give rise to so-called cavitation erosion, whereby the resulting gas bubbles implode due to a sudden increase in the pressure acting on their surfaces behind the flow resistors. The continuous stress acting on the material causes fatigue, and particles are broken away [7]. Flow resistance also gives rise to loud noises, as well as instability in throttle controllers. If the fluid–gas mixture is led back into the tank downstream of a resistor without being pressurized, foam is produced as a result of the low redissolving velocity. Cavitation and cavitation erosion, which presents a serious problem, particularly with respect to water-based fluids, can be reduced by suitable design measures, such as the selection of special materials

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-15

for surfaces which are susceptible to erosion, guiding the cavitation stream into uncritical areas away from the walls, or diminishing the differential pressure at one resistor by fitting several resistors, one behind the other.

11.3.7 Air Release Capacity and Foaming Tendency Air bubbles may be entrained where fluid flows into the tank in an unimpeded stream. Apart from this, air may also get into the fluid as a result of leakages in the system, eddy currents in the tank, or cavitation. This air must be released again on the surface before it can be sucked into the pump. As shown by the following calculation, the air bubbles’ rate of ascent depends on their diameter and on the viscosity and density of the fluid. The bubbles’ lifting force amounts to FA = 43 π(ρFl − ρL ) · r 3 g

(11.16)

where r is the radius of the bubble, ρFl is the density of the fluid, and ρL is the density of the air. According to Stokes, the flow resistance of spherical bodies for very low Reynolds’ numbers is given by FW = 6π ηvr

(11.17)

This means that in the case of equilibrium, the rate of ascent is given by v=

2 r 2g (ρFl − ρL ) 9 η

(11.18)

Fluids like these have a good air release capacity for a given operating viscosity, which allows the undissolved air to coagulate into larger bubbles and therefore rise more quickly. This may be boosted by surface-active additives. Unfavorable currents in the tank may slow the air release process down considerably. Wherever possible, design measures must be implemented to prevent air getting into the fluid and facilitate the release process. ASTM D3427 provides experimental details for determining air release properties of fluids. One negative characteristic of hydraulic fluids is the formation of foam on the surface as a consequence of the released air. This behavior may be prevented by the use of suitable additives, but these also have a detrimental influence on the air release capacity.

11.3.8 Aging Behavior The term “aging” includes changes that take place in the composition and chemical structure of a hydraulic fluid. Aging is brought about by such chemical reactions as oxidation, hydrolysis, polymerization, and thermal decomposition, or by mechanical influences, such as shearing action. Oxidation refers to the reaction with O2 producing left-over acids. Polymerization refers to the enlargement of hydrocarbons resulting from the formation of side chains or macromolecules. This process produces waste products such as sludge or resin-like coatings on components. Hydrolysis refers to the cracking of esters when they come into contact with water. The aging process breaks down or destroys the additives and, at the same time, changes the molecules of the base fluid. It is accelerated by high operating temperatures and contamination through extraneous air, water, and metallic catalysts, predominantly copper, copper alloys, and iron. One example of a measure for the aging condition of a fluid is the acid number (AN). The acid value indicates the acid content of a fluid by defining how many milligram of caustic potash solution would be required to neutralize 1 gram of a sample: AN =

© 2006 by Taylor & Francis Group, LLC

mgKOH gFl

11-16

Handbook of Lubrication and Tribology 0.6 Without contamination 3 vol. % undissolved air 8 vol. % undissolved air 2.5 vol. % water

[mg/g]

TAN

0.4

250 bar 70°C E

0.2

0

FIGURE 11.9

0

250

500 Test period

750

[h]

1000

Influence of air and water on the aging of undoped mineral oil.

The aging stability of a hydraulic fluid can be determined by means of easy laboratory tests, for example, using a Baader device in accordance with DIN 51 554. Results with more practical relevance can be obtained from test-rig experiments [8] as shown in Figure 11.9. The figure shows the way in which air and water influence the increase in the acid value of an unfortified mineral oil over 1000 h under a constant high load. A study of ready-doped hydraulic fluids shows that the TAN drops initially as a result of the breakdown of the acid antiaging additives with a subsequent transition to a progressive ascending gradient. The aging stability of mineral oils, vegetable oils, and HFD fluids can be improved by means of certain additives. As far as water-based fluids are concerned, the base fluid does not age. However, a reduction in the water content must be anticipated during the fluid’s useful life due to evaporation, which results in an increase in the additive concentration.

11.3.9 Material Behavior Hydraulic fluids should not attack metallic materials. Some HFC fluids react aggressively with tin and cadmium. HFD fluids attack aluminum and aluminum alloys in the presence of friction stresses. Far greater difficulties are caused by the ways in which the fluids react with plastics of the types used for seals, hoses, paints, lacquers, and varnishes. Virtually all of the material currently available on the market can be used in conjunction with mineral oils. The only materials which can be used for HFC fluids are silicone rubber and Teflon materials, while only Viton and Teflon materials can be used with HFD fluids. The polar character of environmentally friendly ester fluids leads to a noticeable swelling in conventional standard elastomers. Furthermore, if the compatibility with aged fluids is not known to an adequate extent, a number of adapted materials now exist which permit comparatively safe operation. More serious problems are being encountered with hoses or contamination with water in isolated cases. In addition to this, the products of aging may give rise to problems. Epikote and DD paints are resistant to HEPG, HFC, and HFD fluids to a certain extent. In this respect, it is a good idea to avoid painting the inside surfaces of tanks and to choose suitable materials to avoid the corrosion problem.

11.3.10 Detergent and Dispersant The separation of water and solids in the tank offers a satisfactory means of keeping the fluid clean, at least for low circulation rates. Mineral oils and HFD fluids usually have a good separating capacity, which is referred to as their detergent properties. The precipitation rate of water follows Stokes’ law, analogous

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-17

to air bubbles’ rate of ascent (Equation 11.18): v=

2 r 2g (ρFl − ρw ) 9 η

(11.19)

where ρw is the density of the water. Because water has a higher density than mineral oil, it separates out at the bottom of the tank, but it separates out at the surface of HFD fluid because it has a lower density. Here too, the detergent behavior is boosted by the coagulation of smaller drops into bigger ones. In special cases where there is a very high risk of contamination, water and solids are finely distributed and kept in suspension by dispersant substances. This eliminates the risk of malfunctioning valves and detrimental effects on the antiwear properties and loadbearing capacity. The polar ester fluids have a tendency to form stable suspensions and are also capable of dissolving relatively large quantities of dirt particles and water. However, the hydrolytic effect of water on ester compounds is such that appropriate design measures must be implemented to prevent the ingress of water into the hydraulic system where native and synthetic esters are used.

11.3.11 Thermal Capacity and Thermal Conduction The specific heat and thermal conductivity parameters influence the steady-state temperature in a stationary hydraulic installation, as well as changes in temperature with alternating loads. These parameters are needed for the dimensioning of heat exchangers. Because of the higher specific heat and better thermal conductivity of water-based fluids, the steadystate temperature of systems using these fluids can always be kept lower than the temperature of systems using mineral oils or HFD fluids under otherwise identical conditions. This is also necessary due to the higher vapor pressure of water.

11.3.12 Flammability The flash point is the temperature at which fluid vapor which has developed in a test vessel ignites for the first time when approached by a flame (ISO 2592). It is lower than the inflammation temperature of the fluid. The inflammation temperature is the temperature at which droplets of fluid ignite spontaneously under certain test conditions. This temperature is used as a criterion for the hydraulic fluids being fire resistant. The ignition delay is the period between applying 40 ml of a fluid onto a molten mass of hot aluminum at 800◦ C and the fluid bursting into flames. This period is substantially longer for the fireresistant fluids than it is for mineral oil. This procedure and the designation “ignition delay” are not standardized. Reference 11 provides a summary of a range of fire-resistance tests that may be used to evaluate hydraulic fluid flammability potential.

11.3.13 Biodegradability and Toxicity The biological degradation of a substance involves the use of micro-organisms in an aqueous environment to convert the substance into CO2 and biomass. From the point of view of environmental protection, it is important that degradation takes place quickly and that it is complete. As shown in Figure 11.10, native oils have decomposed by almost 100% after 21 days under very idealized test conditions, whereas mineral oils merely achieve a degradation rate of 25%. Paraffin oils, which do not contain any aromatic compounds, reach a rate of 40%. The bandwidth of synthetic esters is the result of their multiple possible variations. A hydraulic fluid must have a degradation rate of 80% for the environmental symbol (“Blue Angel”) 79. The toxic influence of hydraulic fluids on mammals, plants, and bacteria is essentially determined by the additives. According to the German water resources act (Verwaltungsvorschrift wassergefährdende Stoffe), it is evaluated according to three water pollution classes (WGK 1 to 3): slightly hazardous to water, hazardous to water, and very hazardous to water. Environmentally compatible fluids (vegetable oils and some synthetic esters) may not be regarded as being hazardous to water or are put into the WGK 1 class

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11-18

Handbook of Lubrication and Tribology 100%

WGK (1–3)

FIGURE 11.10

1

1–2

1–2

1–3

Paraffin oils

Mineral oils

Synthetic esters

Polyglycols

Vegetable oils

50%

Biological degradation as per CEC-L-33-T82 (21 days)

1–2

Bandwidths of degradation rates and water pollution classes (WGK) for various base fluids.

according to the additives used. Standards for test procedures of biodegradability and toxicity can be found in ASTM D7044, ISO 9439 and OECD procedures OECD 201, 202, 301B and 301B.

11.4 Fortifying Hydraulic Fluids Native and synthetic fluids, as well as natural hydrocarbon oils, are not always capable of meeting the requirements of modern hydraulic components and systems. The quality of the base fluid can be improved only to a certain extent by modifying the manufacturing processes, which means that additional chemical substances must be used to improve the fluid’s performance. These substances are referred to as additives. The chemical degradation of the additives during the fluid’s life cycle is such that a high-grade base oil with few additives should be given precedence over a lower-grade base oil with a high additive content to ensure a long useful life [1,5]. Additives are classified into those which influence the physical and chemical properties of the base fluids, such as VT behavior, crystallization tendency, and aging stability. At the same time, other additives act on the boundary surface between the fluid and components or impurities and thereby improve the frictional and wear behavior, prevent corrosion, or keep particles in suspension. Hydraulic media are made up of a base fluid, which is doped with a so-called additive package to produce a ready-to-use formulation. The effects brought about by the chemical actions of the various additives may be synergetic or even antagonistic. Many additives perform several functions which reduces the possibility of reciprocal interference by individual additives. This group of substances is referred to as the group of “multipurpose additives.” Hydraulic fluid classifications are provided in ISO 6734/4 and ASTM D6158 and D7044.

11.4.1 Oxidation Inhibitors The oxidation reactions occurring in a hydraulic fluid as a result of atmospheric air at higher temperatures cause the fluid to age. Metal ions, such as copper, iron, and lead, may also exert oxidative or reductive influences and accelerate the aging process [6]. Modern hydraulic fluids require a balanced number of so-called oxidation inhibitors in order to counteract these undesirable effects. At the end of the refining process, base mineral oils contain natural inhibitors in the form of sulfur and nitrogen compounds. Frequently inadequate, the resulting oxidation stability is increased by adding other specific compounds. The most important representatives of this group are sulfur compounds, phosphor compounds in the form of phenol phosphate derivatives, compounds of sulfur and phosphor in the form of zinc-dialkyldithiophosphates, phenol derivatives in the form of sterically hindered polyalkyl phenols, and amines.

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-19

11.4.2 Metal Deactivators If a catalytic acceleration of the oxidation process — more precisely the autoxidation process — in hydraulic media caused by metal ions, and copper and iron in particular, is to be prevented, these ions must be “masked out.” Suitable additives are what are referred to as chelating agents, for example, N -salicylidene ethylene diamine, which are effective in very low concentrations, binding ions in the form of complexes. Film-forming media produce a passivating protective film on metal surfaces, thereby preventing the transfer of catalytic ions to the fluid, as well as the oxidative attack by oxygen and oil aging products on the surface.

11.4.3 Wear Protection A high load-bearing capacity must be given to a hydraulic medium where larger forces are to be transmitted with a low rate of wear. The fluid may be doped with so-called high-pressure or EP (extreme pressure) additives for this. EP additives act on sliding faces which are subjected to high pressure and high temperature loads at the transition between hydrodynamic lubrication and mixed friction. They form metal compounds on the surfaces of the sliding, mating friction faces, which are solid under normal circumstances but liquid to slippy under wear conditions. This prevents the surface from becoming worn. Important EP additives include organic sulfur and phosphor compounds and combinations of these elements. Zinc dithiophosphates (ZnDTP), in particular, are used in a wide range of applications. There is trend toward fluid formulations that are free of heavy metals, however, which means that ZnDTP will play a less significant role in the future. Methods to evaluate wear characteristics of hydraulic fluids are summarized in ASTM D7043, D6158 and D6973. Classification Standards are provided in Sections 11.4 and 11.5.

11.4.4 Friction-Reducing Agents, Friction-Modifiers The mixed friction zone is crossed when sliding metal faces run in and out, which means that mild highpressure additives are used for many applications, to prevent stick-slip or noise and reduce the frictional forces and thereby save energy. Also referred to as “friction modifiers,” these additives generally work by forming thin layers on the mating faces by means of physical adsorption; they comprise polar substances, such as fatty alcohols, fatty acids or fatty-acid esters, amides, or salts. HE media offer sufficient lubricating properties as a result of the polar base fluid.

11.4.5 Viscosity Index Improvers VI improvers are additives that improve the viscosity/temperature behavior of fluids, that is, they diminish the reduction in viscosity as a function of the temperature. The VI improvers usually used today are made up of linear polymer molecules, which are effective in that they increase the viscosity of a fluid to a different degree at different temperatures. This affects the flow behavior of the hydraulic medium in such a way that it can no longer be referred to as a Newtonian fluid. The dynamic viscosity changes with the shear stress. Another very important aspect of VI improvers is that they become more sensitive to mechanical stresses as their molecular mass increases. Shear stresses that occur when the fluid flows through control edges, for example, irreversibly break polymer molecules down into fragments, which ultimately causes a reduction in viscosity at higher temperatures. HE fluids and water-based media have a naturally high VI and do not therefore require any VI improvers.

11.4.6 Setting Point Depressants Where machines and installations are used in locations or regions where the ambient temperature is well below freezing point, the flowability of the medium must be assured during a cold start. As they cool down and reach their solubility limits, mineral-oil-based hydraulic media precipitate n-paraffin hydrocarbons in crystalline form as needles and plates, which form a matted network and stop the oil flowing; that is, it sets.

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11-20

Handbook of Lubrication and Tribology TABLE 11.4

Types and Causes of Contamination in Hydraulic Fluids Types

Particles

Liquids Molecular

Chips, scale, etc. Dust, sand Rubbed-off parts Water Extraneous liquids Sludge, resin, acids Metal ions Oxygen

Causes/sources Production, assembly Storage, installation, maintenance, drawn in during operation Wear Storage, maintenance, drawn in during operation Combined systems, maintenance Aging products Friction, wear Dissolved air

The low-temperature properties of these oils can be improved by means of exhaustive deparaffinization. This is an expensive process, however, so deparaffinization is carried out only to a setting point of around −15◦ C. Further improvements can be achieved by using setting point depressors based on the products or polymerization and condensation. Typical examples of these include polymethacrylates, alkylphenols, and copolymers of vinyl acetate and ethylene.

11.4.7 Antifoaming Agents Heavy foam formation exerts a detrimental effect on the lubricating properties of hydraulic fluids, promotes their oxidation, and can cause air to be sucked into the pump. As far as pure mineral oils are concerned, the stability of the foam is a function of the viscosity and the surface tension. According to Stokes’ law, the speed at which the air bubbles precipitate is proportional to the square of their diameter and inversely proportional to the viscosity. Foam with large bubbles which disappear quickly is produced in a low-viscosity fluid, whereas finely distributed small bubbles form in highly viscous fluids and these make the foam highly stable. Additives derived from liquid silicones (polydissethyl siloxanes in particular) have proven to be the most effective antifoaming agents.

11.4.8 Detergents and Dispersants Owing to their large-scale use in engine oils, the surface-active detergents and dispersants have become the most significant types of additives, accounting for approximately 50% of the market share. Their function is to keep oil-insoluble substances, resinous and bituminous oxidation products, and water in suspension, or to accelerate their sedimentation in order to prevent deposits forming on metal surfaces, thickening of the fluid, precipitation of sludge, and corrosion. The dispersant or detergent effects of many of these additives depend on the respective concentration. Dispersants are ash-free organic compounds which prevent the flocculation or coagulation of colloidal particles. The oil-soluble or finely dispersible metal salts of organic acids known as detergents are reputed to have good dirt-dissolving properties. Both types of additives complement one another with respect to their effective characteristics, which are supplemented by the ability to neutralize acid products, thereby inhibiting oxidation. The triple action of the dispersant, cleaning, and neutralizing substances is such that relatively large quantities of these HD (heavy duty) additives are needed and used. If water gets into hydraulic fluids, relatively stable water-in-oil emulsions with disturbing properties may be produced, which can frequently be de-emulsified only by changing the interfacial surface tension. Basically speaking, all types of surface-active compounds are suitable as de-emulsifying agents. De-emulsifying agents increase the foam formation tendency and may therefore be added only in very low concentrations. Emulsifiers are particularly important in their capacity as emulsifying aids for fire resistant hydraulic fluids. Because of their hydrophobic-hydrophilic molecular structure, emulsifiers have surface-active

© 2006 by Taylor & Francis Group, LLC

Hydraulic Fluids

11-21 Number of particles per 1 ml more than up to

Maximum number of particles per 100 ml in the specified particle size range Class 5–15 mm 15–25 mm 25–50 mm 50–100 mm >100 mm 125 250

22 44

4 8

1 2

0 0

0 0

500

89

16

3

1

1

1,000

178

32

6

1

2

2,000

356

63

11

2

3

4,000 8,000 1,6000

712 1,425 2,850

126 253 506

22 45 90

4 8 16

4 5 6

32,000

5,700

1,012

180

32

7

64,000

11,400

2,025

360

64

8

128,000 256,000

22,800 45,600

4,050 8,100

720 1,440

128 256

9 10

512,000 91,200 1,024,000 182,400

16,200 32,400

2,880 5,760

512 1,024

11 12

Scale number

2,500,000 1,300,000 640,000 320,000 160,000

2,500,000 1,300,000 640,000 320,000

> 28 28 27 26 25

80,000 40,000 20,000 10,000 5,000

160,000 80,000 40,000 20,000 10,000

24 23 22 21 20

2,500 1,300 640 320 160

5,000 2,500 1,300 640 320

19 18 17 16 15

80 40 20 10 5

160 80 40 20 10

14 13 12 11 10

2.5 1.3 0.64 0.32 0.16

5 2.5 1.3 0.64 0.32

9 8 7 6 5

0.08 0.04 0.02 0.01 0.00

0.16 0.08 0.04 0.02 0.01

4 3 2 1 4 mm

1452.53

18

>4 mm

1452.53

>6 mm

274.4

15

>6 mm

274.4

>14 mm

18.51

11

>14 mm

18.51

ISO 4406:1999 18/15/11

Count result Particle size >5 mm >15 mm

Number per 1 ml 186.2 14.45

>25 mm

4.9

>50 mm

0.72

>100 mm

0.08

Particle range

Number per 100 ml

Class

5–15 mm

17,175

7

15–25 mm

955

5

25–50 mm

418

6

50–100 mm

64

6

>100 mm

8

5

NAS 1638 Class 7

Example for classification of a count result

Recommendations for degrees of purity Type of hydraulic system

Degree of purity

Sensitive control systems, laboratory, aerospace systems

11/8

High-performance servo systems, high-pressure systems with long life

14/11

Reliable, high-quality systems, general machines

16/13

Medium pressure range, general machines + automotive

18/14

Low pressure range, general/heavy-duty machines, automotive

19/15

Low-pressure systems with large tolerances

21/17

FIGURE 11.11 Degrees of purity for hydraulic fluids as per ISO 4406 and NAS 1638.

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Handbook of Lubrication and Tribology

properties and reduce the interfacial surface tension of the water to facilitate the formation and stability of the emulsion. A distinction is made between anionic, cationic, and nonionic emulsifiers.

11.4.9 Corrosion Inhibitors Corrosion occurs when a metal surface is exposed to oxygen (or another aggressive substance) and moisture at the same time. The corrosion caused by electrolytic processes can be prevented to a great extent by the formation of a nonmetallic protective layer. Effective inhibitors should adhere firmly to the metal and produce a film that is impermeable to water and oxygen. Nitrogen compounds, fatty-acid amides, phosphoric acid derivatives, sulphonic acids, sulfur compounds, and carboxylic acid derivatives are particularly important in this respect.

11.5 Impurities in Hydraulic Fluids There are many sources of contamination or impurities in hydraulic fluids. Refer to Table 11.4. These are already found during the manufacturing and assembly processes in the form of metal chips, grinding dust, welding beads, sand, scale, etc. The initial contamination of freshly supplied fluid is often substantially greater than is permitted for normal operation and this can increase further if the fluid is not stored properly. Under normal operating conditions, dust, fine sand, condensation water, and rainwater from the environment are drawn into the tank through the air vent and into the system by means of deposits on the piston rod. Rubbed-off metal parts are found in the system in the form of particles and released metal ions, along with rubbed-off parts of seals as a consequence of wear. The chemical change brought about in the fluid by temperature, pressure, and shear stresses leads to aging products, such as sludge, resins, and acids, which are usually caused by oxidation with the dissolved atmospheric air. Contamination not only affects the aging of the medium but also the useful lives and functions of the components. Initiated by the introduced energy, several contaminants frequently act together, which means, for instance, that water, oxidized oil molecules, metal ions, and used additives combine to produce sludge. Dirt particles are transported to all parts of a system with the fluid. They may lead to a direct failure if particles cause the slide in a valve to jam, for example, or block control nozzles. A far more significant influence is brought about by the impurities; however, additional wear in the components. Abrasion is caused by particles caught between mating faces, and erosion takes place on edges and surfaces which are exposed to the fluid flowing at high velocity. Solid contaminants are described according to purity classes, which define the maximum permitted number of particles of a particular size. Recommendations are given for the applicable purity class according to the components used, Figure 11.11. In this respect, apart from the components’ sensitivity to wear, the gap widths specified for the design and the control nozzle diameter also play particularly important roles. Continuous maintenance of the fluid by filtering and separating water out, as well as preventive measures, such as thorough cleaning of the components during the assembly process and using filter systems TABLE 11.5

Characteristic Values of Hydraulic Fluids

Density at 15◦ C (g/cm3 )

Kinematic viscosity at 40◦ C (mm2 /sec) Mean compression modulus E (N/m2 ) Viscosity/temperature index Specific heat at 20◦ C (kJ/kgK)

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HLP

HFA (3%)

HFD

HETG

0.87 10 to 100 2 × 109 100 2.1

1.0 0.7 2.5 × 109 — 4.2

1.15 15 to 70 2.3–2.8 × 109 30 LWI 10

30 Pass >10

30 Pass >10

30 Pass >10

Note: Additive solubility — must be filterable to 3 µm (beta = 200 filter rating) without loss of additive(s).

capable of penetrating into the body of wire rope, yet maintain retarded dripping qualities for operation over wide temperature ranges. These products shall contain chemical EP and solid film additives to improve film strength and control fretting and rubbing friction during operation.

20.7.5 Antiwear Hydraulic Oil This specification covers premium circulating oils produced from refined mineral oil-base stocks, and compounded with antiwear additives for high load-carrying ability. These materials are primarily intended for use in hydraulic systems operating within an ambient temperature range of −18◦ C/0◦ F to 54◦ C/130◦ F. They may also be used to lubricate high-speed plain or rolling element bearings, lightly loaded enclosed gear drives, and miscellaneous items such as links, pins, and bushings operating in circulating, sump (splash), or total loss applications. Typical manufacturers’ performance requirements are given in Table 20.15.

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Handbook of Lubrication and Tribology

References [1] Anon, The World of SKF, SKF Publication 3277E Reg 0770.75.000.19803, Sweden, p. 202. [2] Lehner, S. and Jacobs, G., Contamination sensitivity of hydraulic pumps and valves, Tribology of Hydraulic Pump Testing, ASTM STP 1310, George E. Totten, Gary H. Kling, and Donald J. Smolenski (eds), American Society for Testing and Materials, 1996. [3] Tessman, R.K. and Hong, I.T., Hydraulic pump contaminant wear, Hydraulic Failure Analysis: Fluids, Components and System Effects, ASTM STP 1339, G.E. Totten, D.K. Wills, and D. Feldman (eds), American Society for Testing and Materials, West Conshohocken, PA, 2001. [4] Roberts, W.H., The lubrication of rolling bearings, 3rd Annual Conference on Industrial Tribology, Caulfield Institute of Technology, Melbourne, Australia, 48, 1982. [5] Bensch, L.E., Fitch, E.C., and Tessman, R.K., Contamination Control for the Fluid Power Industry, Pacific Scientific Company, Montclair, CA, 1978. [6] Jagger, E.T., Lip seals, Tribology Handbook, 2nd ed., M.J. Neale, (ed.), Newnes-Butterworth, 1975, A35. [7] Lawrence, R.T., Soft piston seals, Tribology Handbook, 2nd ed., M.J. Neale, (ed.), London, 1975, A37. [8] Anon, SKF Electronic Handbook, SKF Publication 4485E 10000, Sweden, 1995. [9] Anon, FAG Technical Publication and General Catalogue, Compact Disc, FAG Australia Pvt. Ltd., 2000. [10] Ioannides, E. and Jacobson, B., “Dirty lubricants — reduced bearing life,” Ball Bearing Journal, Special Issue 89, 22, 1989. [11] Anon, The Lubrication of Rolling Bearings, Publ. No. WL 81 115/2 EF/98/10/87, FAG Australia Pvt. Ltd. [12] Harris, J.H., The Lubrication of Rolling Bearings, Shell-Mex and B.P. Ltd., London, 1967. [13] Cheng, H.S., Elastohydrodynamic Lubrication, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 1983, 139. [14] Anon. Roller Bearing Lubrication Handbook, Fuchs Petrolub AG Oel + Chemie, GfT Work Sheet 3, 1994. [15] Ribble, H.C., Cast Bronze Bearing Design Manual, 2nd ed., Cast Bronze Bearing Institute Inc., Cleveland, Ohio, 1965. [16] Wills, J.G., Lubrication Fundamentals, Mobile Oil Corporation, Marcel Dekker Inc., New York, 1980, 112. [17] Scott, W., Design detail affecting reliability, Presented at International Mechanical Congress, MECH ’91 I.E. Aust., Sydney, Australia, 1991. [18] AGMA Specifications, Lubrication of Industrial Open Gearing. (AGMA 250.02), The American Gear Manufacturers Association, Washington, D.C. [19] Klaus, E.E. and Tewksbury, E.J., Liquid lubricants, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 229, 1983. [20] Wellauer, E.J. and Holloway, G.A., Application of EHD oil film theory to industrial gear drives, Trans. ASME, J. Eng. Ind., 98B, 626, 1976. [21] Roth, H.E., Design and manufacture for load distribution, Presented at International Conference on Mining Machinery, I.E. Aust, National Conference Publ. No. 79/5, Brisbane, Australia, 368, 1979. [22] Annual Book of ASTM Standards, Section 5, Petroleum products, lubricants and fossil fuels, Vol. 05.03 Petroleum Products and Lubricants (III), D4636-latest; Catalysts, 1997. [23] Williamson, J.B.P., The shape of surfaces, Handbook of Lubrication, Volume II, Theory and Design, E.R. Booser (ed.), CRC Press, Boca Raton, FL, 3, 1983. [24] Scott, W. and Hargreaves, D.J., Specifying surface roughness for spur and helical gears, Tribology for Energy Conservation, D. Dowson et al. (eds), Elsevier Science B.V. 267, 1998. [25] Scott, W., Report on the Lubrication of Dragline Gears for BMA, Scott Tribology Services Pty. Ltd., Brisbane, Australia, August 2004.

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Mining Industry

20-25

[26] Bartz, W.J. and Kruger, V., Test Method for Evaluating the Influence of Lubricants and Lubricant Additives on Fatigue Failure of Quenched and Tempered Case-Hardened Spur Gears, Rolling Contact Fatigue-Performance Testing of Lubricants, R. Tourret and E.P. Wright (eds), Institute of Petroleum, London, 161, 1977. [27] Anon, Hoist rope lubrication criteria, Battelle Columbus Laboratories, Report No. PB80-182959, Prepared for Bureau of Mines, Washington, D.C., 1978. [28] Critchlow, J.P. and Flynn, R.W., Wire rope lubricants and lubrication, Lubrication Engineering, August 1951, 178–181 and 195. [29] Kaderjak, G., Stranding head for the internal lubrication of steel wire ropes, Wire World International, 18, 35, Jan–Feb 1976.

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21 Farm and Construction Equipment 21.1 21.2 21.3 21.4 21.5

Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Types of Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Development of Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . Know the Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-2 21-2 21-2 21-3 21-3

Diesel • Gasoline • Liquid Propane (LP) Gas • All Engines

21.6 Clutches, Fluid Couplings, and Converters . . . . . . . . . .

21-4

Clutches • Fluid Couplings • Converters

21.7 Electric Motors, Generators, and Rectifiers . . . . . . . . . .

21-5

Electric Motors, Generators • Motors, Generators • Rectifiers

21.8 Gear Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-6

Powershift and Automatic Transmissions • Manual Transmissions, Enclosed Gear Drives • Open Gear Drives

21.9 21.10 21.11 21.12 21.13 21.14

Couplings. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chain Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Wire Ropes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Crawler Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydrostatic Hydraulic Systems . . . . . . . . . . . . . . . . . . . . . . . Bearings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-9 21-9 21-10 21-10 21-10 21-11

Antifriction Bearings • Plain Bearings • Bearing Lubricants

R. Lal Kushwaha Professor, Machinery Systems Ag. and Biosource Eng. Dept. University of Saskatchewan

Jude Liu Post-Doctoral Fellow Ag. and Biosource Eng. Dept. University of Saskatchewan

21.15 Pneumatic Equipment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-13

Air Compressors • Air Cylinders • Air Motors • Pneumatic Tools

21.16 Good Lubrication Practice . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-14

Management • Operators • Maintenance • Supplier • Oil Analysis

Acknowledgment. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21-18 21-18

21-1

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Handbook of Lubrication and Tribology

21.1 Introduction Farmers and contractors in the United States currently claim assets of close to $2 trillion and employ over 8 million workers. The value of farm cash receipts exceeds $240 billion while the value of construction is well over a trillion dollars per year. The equipment currently consumes over 30 billion gal of fuel per year and uses more than 300 million gal per year of oils and greases (www.nass.usda.gov/census/ and www.census.gov/mcd/). Easily, farming and construction together are this country’s biggest business and the world’s largest and most important endeavor, from providing the four Rs of road work (reconstruction, resurfacing, rehabilitation, and recycling) to producing food and fiber. Their machines are built for utility and the ultimate use of raw power, but are still beautiful in their complexity. Any farm tractor can be fitted with a dozer blade, bucket, or backhoe to move earth for any construction project on the spread and the farm equipment manufacturers have naturally progressed to the manufacture of heavier units for construction of buildings, pipelines, and roads. Whereas each industry has machines particular to certain work, all employ the same basic power and drive train systems, all work off the road, and all are subject to the same heavy-duty service and inhospitable outdoor working conditions.

21.2 Types of Equipment There are more than 100 distinctive equipment types, but they can be put into a few basic categories according to the work they do: • Digging, filling, earth-moving Backhoes, shovels, scrapers, excavators, dozers, draglines, tractors, trenchers, and graders • Loading Loaders, shovels, excavators, and lift trucks • Hauling Trucks and trailers • Drilling, breaking Rippers, drills, hammers, and planers • Processing Harvesters, crushers, screens, and mixers (asphalt, concrete) • Placing, laying Cranes, pipelayers, layers (asphalt, concrete), and rollers • Auxiliary Air compressors, pumps, pneumatic tools, welders, finishing tools, and generating plants Since attachments can be made to many of the basic machines, most do more than one job. If the changes could be made overnight, a crane could be digging one day, driving piles the next, and placing steel the day after that. An excavator could be digging a basement one day and punching holes in old concrete the next. A typical farm tractor with a variety of add-ons can plow, load and haul, or dig a trench.

21.3 Development of Lubricants The agencies involved in engine oil classifications: 1. 2. 3. 4. 5. 6. 7.

American Automobile Manufacturers Association (AAMA) Japan Automobile Manufacturers Association (JAMA) International Lubricant and Standardization and Approval Committee Society of Automotive Engineers (SAE International) American Society for Testing and Materials (ASTM) American Petroleum Institute (API) Federal Government — Environmental Protection Agency (EPA)

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Farm and Construction Equipment

21-3

TABLE 21.1 API Oil Classifications Year

Light-duty engine oil

Heavy-duty engine oil

1955 1968 1972 1980 1985 1988 1991 1993 1994 1995 1996 1998 2001 2002 2004

— SD SE SF — SG — SH — — GF-2 — GF-3 — GF-4

CD — — — CD-II CE CF-4 — CF, CF-II CG-4 — CH-4 — PC-9 —

Table 21.1 lists the development of various lubricants classifications according to API for the light-duty engines (spark ignition) and the heavy-duty engines (compression ignition) (Imperial oil, 1999).

21.4 Know the Machine Good lubrication practice requires a thorough knowledge of the machine, from the prime mover through to the final application of the force to get a job done. Stationary equipment has essentially one power flow, to crush rock, size aggregate, convey materials, etc. Mobile equipment incorporates two power trains — one to propel the machine, another to do the work intended. A careful trace of the flow of power in the machine is vital, not only to recommend the proper lubricants, but also to assure that no elements in the system are missed. Modern machinery employs every conceivable method of transmitting power to the load or the road. Electric motors or internal combustion engines exert power through transmissions or gear boxes to final drives, or through converters to chain and gear drives, or send power to do work through hydrostatic pump and motor systems. Loads are handled by wire ropes, hydraulic cylinders, and gear-cases. Power is transmitted through couplings and shafts. All machine makers provide operator’s manuals with a lubrication chart; if not for the composite machine, at least for the various components. Many of these unfortunately become outdated or get lost. It falls on the lubrication engineer, even if a metal plate or a decal is attached to the machine, to use his experience, knowledge of power transmission, and familiarity with lubricants and lubrication to see that best lubrication practice is performed.

21.5 Engines 21.5.1 Diesel Most farm and construction machines are powered by diesel engines since this type of work calls for low-speed, high-torque operation, and lower rates of acceleration than found in many gasoline or gas (Otto cycle) engine applications. The diesel combustion cycle is well fitted for this type of service with high-torque rise at lower engine speeds being part of the basic design. The engines may be of two- or four-stroke cycle, direct injected or prechamber, turbocharged or naturally aspirated, and air or water cooled. They are a rugged and dependable power source, operate many hours before major overhaul is required, and cost per brake horsepower–hour is usually lower than for other engines.

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Handbook of Lubrication and Tribology

Because of the high compression ratio of diesel engines, 14 to 20 : 1, protection of the piston ring belt area from deposits is of prime importance. As a general rule, engine oils of American Petroleum Institute (API) Service Classification suitable for naturally aspirated engines and for supercharged or turbocharged engines should be used. From the environmental viewpoint, it is important to match the acid neutralizing capability of the oil, indicated by total base number (TBN), to the fuel used.

21.5.2 Gasoline These engines are used in smaller horsepower ranges for mobile and stationary equipment and may also be of the two- or four-stroke cycle design, air or water cooled. In general, more appropriate oils of API Service Classification suitable for four stroke engines as well as for two-stroke engines (oil premixed with the fuel) should be used. Some engine manufacturers may require the use of very low ash content oils (≈40 ppm), specifically formulated for this type engine.

21.5.3 Liquid Propane (LP) Gas Engines in lift trucks and smaller unit applications may be of this design, using fuel commonly called “propane.” This fuel burns cleanly and causes little or no crankcase oil dilution, so oils with mild detergentdispersancy and with oxidation and bearing corrosion inhibitors may be used in most cases. To offset increasing viscosity due to oxidation, make-up oil of the next lower SAE number may be required.

21.5.4 All Engines While, in general, the service (“S”) classifications of engine oils are used to lubricate gasoline and LP gas engines, and commercial (“C”) classifications are recommended for diesel engines, all engine manufacturers recognize the specific ASTM test requirements of all classifications and have experience with particular formulations. They seek to incorporate in their specific recommendations the properties required by their engines for their design purposes. Therefore, most manufacturers will cross classify their recommendations for engine oil. Many engines are also critical as to oil ash content and additive chemistry. To prolong engine life, most manufacturers will specify the ash percentage (as a maximum, range, or minimum) and require, or effectively dictate, the percentage of certain detergent-dispersant and antiwear additives. Oil change intervals vary with many parameters: engine application, climatic conditions, degree of turbocharging, oil quality, fuel quality, type of oil filtration, fuel consumption, oil consumption, and crankcase capacity.

21.6 Clutches, Fluid Couplings, and Converters 21.6.1 Clutches Conventional dry face type or friction clutches may require lubrication of the pilot and release bearings. Most require the use of high-temperature grease, sparingly applied, but applied often enough. Wet type clutches run in oil which is usually the transmission fluid, may be the engine oil, or may be a separate fluid. Lubricant should be changed periodically.

21.6.2 Fluid Couplings These units absorb the shock loading forces between engine and drive, using oil as the “cushion” and connection between the rotors. The engine oil or other recommended fluid requires regular changing.

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Farm and Construction Equipment

21-5 Boom hoist brakes

Worm gear

Boom hoist Boom hoist drum

Boom hoist drive shaft Drum drive-floating on shaft Chain drives Horizontal independent swing shaft Drum shaft Main drive shaft Swing converter drive

Slide pinion shaft Vertical travel shaft

Hoist converter drive Vertical independent swing shaft Swing shaft Ring gear & roller path Horizontal travel shaft Crawler drive sprocket

FIGURE 21.1 Work and propel power trains showing variety of machine elements to rotate wire rope drums, swing the upper works, and propel the lower works of a crane. (Courtesy of Manitowoc Engineering Company, WI.)

21.6.3 Converters These units have the feature of multiplying engine torque and provide greater flexibility in power train design and operation. In the propel power train of mobile machines they may or may not be integral with the transmission and, in most cases, will automatically lock up into direct drive at specified higher revolutions per minute to increase speed and fuel efficiency. In many machines, for example, cranes, the load power train will find torque converters as separate, controlled units ahead of chain or gear drives (Figure 21.1). Depending on the maker and whether of single or multiple stage design, these units list a variety of oils that should be used. In one type of installation, diesel fuel is the hydraulic medium, the supply and return being piped to the fuel storage tank. Most units, however, either use inhibited oils of various viscosities specially formulated for the purpose, or engine oils and automatic transmission fluids. Since efficiency is affected by viscosity, the manufacturer’s and machine maker’s recommendations should be strictly followed. Converters, along with fluid couplings, are examples of the hydrodynamic principle of fluid in motion.

21.7 Electric Motors, Generators, and Rectifiers 21.7.1 Electric Motors, Generators Some of the largest equipment employs electric motors, both as the prime mover to propel the machine and as the source of power to do the work. Generators may be driven by electric motors or internal

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Handbook of Lubrication and Tribology

combustion engines. Prime examples are shovels, cranes, pumps, crushing and screening plants, and the largest loaders and haulage trucks. In the latter examples, a diesel engine is the prime mover for a generator, with electric motors supplying the final application of power to the wheels.

21.7.2 Motors, Generators These rotate on ball or roller bearings which are generally grease lubricated. In some applications, particularly where thrust bearings are necessary, various viscosities of inhibited oils are recommended. Greases should be sparingly applied and at relatively long intervals. The bearing housing should be maintained at one-quarter to one-third full. Since overfilling causes heat due to internal friction in the grease, the usual procedure is to remove the relief plug, apply grease until new grease is evident, then run the unit for a few minutes to expel excess lubricant. Following the manufacturer’s recommendations is mandatory. Greases should be of high quality, with dropping points suitably high for the operating temperature, exhibiting high degrees of oxidation resistance and water tolerance, and having excellent corrosion protection properties. Another determining factor in the grease formulation is the insulation temperature class of the unit, and this should be known before recommending the lubricant to be used. Periodically, the bearings should be removed, cleaned, and inspected, depending on the severity of service.

21.7.3 Rectifiers These units change AC to DC for the final power source. Transformer oils are used for cooling and require replacement at dictated intervals.

21.8 Gear Drives In both the propel and load power trains, gear drives provide intermediate and final applications of desired torque and speed. Drive trains consist of transmissions, transfer cases, differentials, and final drives, or planetaries (Figures 21.2–21.4). Load trains incorporate transmissions, power take-offs, pump drives, reduction gear cases, and open gears (see Figure 21.1 and Figure 21.5). Gear designs used are spur, spiral bevel, bevel, helical, hypoid, herringbone, and worm. Planetary types of gear drives are used in both the propel and load trains.

21.8.1 Powershift and Automatic Transmissions These units have contributed much to lessening operator fatigue by reducing the need to use the clutch foot pedal and to the life of machine components by the easier selection of the proper gear for the job at hand. In the automatic, gear ratios are selected automatically by the transmission, sensing the proper ratio needed, whereas in the powershift the ratios are selected by the operator moving a lever. Many newer units incorporate the features of both designs. In all these types of transmissions, internal clutches, applied hydraulically, route the power through the proper gear sets to the output shaft. Manufacturers of these transmissions have definite recommendations on lubricants since additive treatment plays a large role in the life and efficiency of the components. Additive systems must be compatible with the frictional material used and maintain the friction-retention properties designed into the clutch packs for the amplitude of torque applied. Fluids of the Type A, Suffix A; DEXRON®; DEXRON®-II and Type C-1 and C-2 specifications are the generally recommended lubricants for automatic and hydraulic transmissions, with provisions by each manufacturer that compatibility and friction-retention properties must be met (refer to SAE J311 Standard). Viscosity characteristics at various temperatures are a consideration in applying the proper lubricant. The last decade has seen the development of many farm and construction wheeled tractors which simplify lubrication by the use of the same oil (in a common sump) for the lubrication of the gears

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Planetary final drive

Differential Drive shaft

Transmission

FIGURE 21.2 The power train of John Deere 9020 Series Wheel Tractors. (Courtesy of John Deere, Deere and Company, Moline, IL.)

Hydraulic motor Drive sprocket Reduction gears

FIGURE 21.3 Drive train showing hydraulic motor, reduction gears, and drive sprocket of a crawler (hydrostatic drive). (Courtesy of FMC Corporation, Cedar Rapids, IA.)

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Reduction gears Clutch

Drive shaft Transmission oil

FIGURE 21.4 The Powershift transmission of John Deere 8020 Series Tractors. (Courtesy of John Deere, Deere and Company, Moline, IL.)

Hydraulic pump drive Hydraulic motors

Hydraulic pumps Swing gear

FIGURE 21.5 Pump drive, pumps for propel and tool hydraulic systems, swing pump and swing motors on a hydrostatic crawler-excavator. (Courtesy of FMC Corporation, Cedar Rapids, IA.)

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in the transmission, differential, and planetaries, and also as the hydraulic medium to operate the tools (Figure 21.4). These lubricants, commonly called “tractor hydraulic fluids” or “tractor universal fluids,” incorporate additives to satisfy the antiwear and lubricity requirements of gears, wet brakes, and, hydraulic pumps.

21.8.2 Manual Transmissions, Enclosed Gear Drives Because of the heavy-duty nature of farming and construction work, gear teeth must be protected by antiwear or extreme-pressure type lubricants. In some designs, however, where increased loadcarrying capacity is not needed,“regular type”lubricants are satisfactory. Recommendations range from API Service Classification GL-l through GL-6 for the gears of drive trains and load trains on mobile equipment to the American Gear Manufacturers’ Association (AGMA) lubricants of rust and oxidation inhibited, fatty oil compounded, or mild extreme-pressure types for some reduction units on stationary machinery (refer to SAE J306 and SAE J308 Standards). Many machine manufacturers allow the use of engine oils in some gear boxes, the antiwear or extreme-pressure quality being sufficient for these designs. In many steering gear boxes and some planetary type gear cases, greases of low to medium consistencies and containing extreme-pressure additives or friction-reducing solids are recommended.

21.8.3 Open Gear Drives This type of power transmission is found on most mobile and stationary equipment. Design is usually of the spur gear type and in most cases calls for manual application of heavy, “residual type” lubricants of seasonal consistency, but usually needing heat to apply. Solvent cut-back types, wherein a solvent eases the application and then evaporates to leave a durable film, are the more convenient lubricants. Some designers call for the additional protection of extreme-pressure additives for their applications. In many cases, the lubricant is applied from a remote-point automatic system, in which case the lubricant must be able to be pumped through long lines. Many machine manufacturers find extreme-pressure greases or those made with heavier oils to be satisfactory lubricants.

21.9 Couplings Many drivelines employ a coupling between the power source and the gear drive to protect against shock loading and allow for possible misalignment of the major units. Those with elastomeric inserts or of flexible plate type require no lubrication, but many couplings are of the grid spring, chain, or gear design and require periodic lubrication with grease or heavy, residual oils depending on the type and conditions of service. Much work has been recently completed on the properties needed by greases. 1. High-speed service — low oil separation, high base oil viscosity at 40◦ C (104◦ F), high dropping point. 2. High-torque, high misalignment (low-speed) service — low oil separation, good lubricity, low dropping point. The latter type of service will be the most likely case in farm and construction machinery, but both types of service will be encountered.

21.10 Chain Drives This type of power transmission is used extensively in off-the-road equipment. Rapid wear will occur if the internal and external parts are not properly lubricated and protected from contamination and corrosion. Larger open chains, such as crawler drive chains, are lubricated manually and usually with engine or gear oils of proper viscosity, but many other applications are satisfied with open gear type lubricants or with

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greases. Such open chains, particularly where exposed to excessive contamination, may be better run with frequent removal for cleaning in fuel oil and soaking in lubricant. Even under ideal conditions, lubricated chains should receive the cleaning and soaking treatment periodically. Enclosed chain drives, being built to greater precision, may be designed to run in an oil bath or be lubricated by drip oilers. Lubricants recommended are seasonal viscosities of engine, extreme-pressure gear, or antiwear oils. To be found on some machines are grease lubricated chain cases, the lubricant being extreme-pressure greases of semifluid or low consistency numbers.

21.11 Wire Ropes These are the prime tools in the construction industry for operating a tool at a distance from the machine, such as in drag line operation or in placing steel for tall buildings. Ropes require not just protection from rust and corrosion, but also wear protection of the individual strands as they rub against each other during flexing over sheaves and around drums. Regular lubrication is necessary because of the continual squeezing out of the lubricant from the inside strands to the outer. Application of lubricants by brush or swab while on the machine, while protecting outer strands and sealing in the inner lubricant, does not effect proper penetration to the inner wearing surfaces. Periodically wire rope should be removed from the machine, cleaned in solvent or kerosene, and then relubricated by passing it through a device which will bend the rope over sheaves in a heated bath or by soaking it therein. Moving cables should receive the most attention, but all ropes need frequent applications of lubricant. One exception is dry rope used in such jobs as drag line operation, where abrasives will tend to be collected by the lubricant. Depending on the operation, a variety of lubricants are called for. The heavy, residual type compounds which may be applied at ambient temperatures or may be heated are the most popular recommendations. These may be solvent cut-back and may also contain extreme-pressure agents. In some dusty atmospheres, lighter-bodied oils better serve the purpose. Gaining wider use are specially formulated grease types, compounded with inhibitors and an effective penetrant.

21.12 Crawler Mechanisms Track laying assemblies or crawlers are used extensively to provide propulsion and tractive efforts for tractors, shovels, cranes, etc. used in construction, with still a few being found on the farm. The mechanism consists of an endless steel belt, driving sprockets and idler tumblers, carrier rollers, and the weightsupporting track rollers. Older types require frequent lubrication with special equipment and special low consistency greases of the tacky, water-resistant type, incorporating a heavy oil. Under some conditions, appropriate viscosities of engine oil or extreme-pressure gear oil are used. Newer designs may also require special equipment for the application of greases or engine oils. Lubrication intervals vary, depending on contaminating conditions and seal design. Some units are designed to hold lubricant until overhaul; others call for daily relubrication to aid in flushing out contaminants. When traveling distances during operation, some of the latter should be greased every 0.8 km ( 12 mi) or 12 h.

21.13 Hydrostatic Hydraulic Systems Fluid power is transferred to mechanical action in farm and construction equipment by two methods: inline by means of a cylinder and piston, or rotary by means of a rotating motor (see Figures 21.2–21.5). The flexibility of the system lies in the prime power being routed to the load through lines rather than through gear trains or wire ropes; power is smoothly applied with infinite control of speed and torque. In the open loop hydraulic system the fluid is pumped from a reservoir by a vane, piston, or gear type pump through lines containing pressure and flow control valves to cylinders or motors that perform the mechanical work, then back to the reservoir. In many systems, pressure and flow are controlled by variable volume/pressure

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compensated pumps and speed is controlled by a variable motor. Many machines utilize closed loop systems, wherein oil from the discharge port of the pump flows to the motor and then directly back to the pump inlet port. This arrangement is called a hydrostatic transmission. Both open and closed loop designs may be found on the same machine. The most readily identified hydraulic system on any machine uses cylinders and pistons for straight-line back and forth motion. These systems are found mainly in the work train to manipulate tools such as the backhoe or bucket, and to operate controls such as those for clutches and brakes. In the propel train, the steering of many wheeled vehicles is done by cylinders, either as an assist or fulltime. Those hydraulic systems employing rotating motors to do the mechanical work are being used increasingly in both the work and propel trains, and in both wheel and track vehicles. On a single machine, for example, some motors may power wire rope drums while others travel the lower and others swing the upper works. Hydraulic motors are of the piston, vane, or gear design and usually (but not always) match the pump as to type. Oil recommendations run the gamut of types produced, depending on the application: mineral oils (nonantiwear and antiwear), water-in-oil invert emulsions, oil-in-water emulsions, water–glycols, phosphate esters and blends, synthesized hydrocarbons, engine oils and the transmission fluids such as General Motors’ Type A, Suffix A, DEXRON®, and DEXRON®-II; Ford’s M2C 33F; Allison’s C-3; and the so-called tractor fluids. With most modern hydraulic systems operating well above 14,000 kPa (2,000 psi), most recommendations will be for those fluids fortified with appropriate degrees of antiwear. Because of the variety of recommendations and the alternates allowed, a farmer or contractor with a large variety of equipment must pay particular attention to hydraulic oil stock and application. Since viscosity is the most important single property, makers of equipment and hydraulic system components recommend use of specific viscosities and viscosity indexes to obtain the best efficiency at the operating temperature of the fluid. Limits take many forms and, since the operating temperature may vary from one point in the system to another, are compromises. Machine makers usually recommend a certain viscosity number for a range of temperatures, some indicating that the lower temperature is the minimum ambient for start-up and the higher one the maximum oil operating temperature. Component manufacturers (and some machine makers) recommend a minimum viscosity index (usually 90), a maximum viscosity at cold start-up, a viscosity range at operating temperature of the fluid, and an optimum operating viscosity. Recommendations vary according to the experience and experimentation of each equipment and component manufacturer for each pump, motor, and cylinder design, so it is paramount that the users of hydraulic fluids be familiar with the requirements of the particular system. In general, hydraulic fluid must have the viscosity (and film strength) to adequately lubricate the closely machined parts of the system, yet not be too high in viscosity to lower efficiency or cause cavitation in the pump (and resulting noise), nor yet too light to lower efficiency or promote leakages. Very few mobile equipment hydraulic systems using mineral oils operate in the ideal range of 54 to 60◦ C (130 to 140◦ F), most of them reaching temperatures of 82◦ C (180◦ F) or even 100◦ C (212◦ F). Water-containing fluids generally should be maintained at operating temperatures below 49◦ C (120◦ F), but in many sealed systems such fluids are used at temperatures up to 82◦ C (180◦ F).

21.14 Bearings Because of the variety of stationary and rolling equipment in the farm and construction industries, every type of bearing design is used to support rotating parts in place and provide their free and efficient motion. While most bearings in enclosed units such as engines and gear boxes receive lubrication by splash, mist, or an internal circulating system, even these units may have bearings that need periodic special attention. Some gear reducers employ an upper bearing that does not receive lubrication by splash or carry-up, and so needs periodic application of grease. Many older engines call for regular attention to accessories such as starter motors, generators, distributors, fan drives, and water pumps. Engine accessories on many newer engines are equipped with nonrelubricatable bearings.

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Wheel bearings, drive lines, and steering and suspension systems of mobile farm and construction equipment require much more frequent lubrication than on-road vehicles. While wheel bearings and many bearings in the work train are repacked or run in an oil bath, most bearings are equipped with fittings or oil or grease cups for daily, weekly, and monthly applications. However, many bearings that demand fairly frequent lubrication have threaded plugs in their housings — these are to be removed and a fitting installed. It is necessary to become familiar with this practice by knowing the drive or load trains of each machine. Not to be overlooked are those moving points such as linkages and pivots that require hand oiling.

21.14.1 Antifriction Bearings Best lubricant recommendation and application for rolling bearings requires knowing the bearing operating temperature and speed. Generally, for operating temperatures below 93◦ C (200◦ F) and speeds below 3000 rpm, the usual greases are used. Above these figures, oils are generally used and oil circulation or mist is considered. However, many of the complex soap, polymer, or inorganic thickened greases perform well up to about 204◦ C (400◦ F) with synthetic formulations being recommended for up to 246◦ C (475◦ F). The caveat in all grease applications is to use greases at maximum operating temperatures of 56◦ C (100◦ F) below the grease dropping point. The operating temperature of the lubricant itself will normally be between 3 and 11◦ C (5 and ◦ 20 F) higher than the housing temperature, which can be fairly easily determined with a contact thermometer. The required viscosity of the oil, or oil component in a grease, can be determined from viscosity–temperature charts for that particular oil. When the bearing load is unknown, the following may serve as a guide for minimum viscosity at the lubricant operating temperature: Bearing type Ball, cylindrical roller, needle Spherical roller Thrust

Minimum oil viscosity (cSt) 13.00 20.45 31.85

Bearing housings should not be overfilled, because overheating may result due to churning of the lubricant. The rule for oil bath housings calls for the lubricant level to be no higher than the center of the lowermost ball or roller when the bearing is at rest; grease lubricated bearings are satisfied with the bearing or bearing housing one-third to half full, keeping in mind that the higher the speed the lesser grease quantity, so long as a sufficient amount is present. On many bearings lubricated through fittings, a vent plug may be provided. This control serves to prevent overfilling since grease is applied during running, with the plug removed until equilibrium occurs. Vent plugs (and pressure relief type fittings) also prevent the blowing out of seals during pressure gun greasing. Many drive line universal joints and other applications not having vents require the use of hand guns to prevent seal or bearing damage.

21.14.2 Plain Bearings Bushings used in farm and construction equipment are not sealed as well as antifriction bearings, and rely on grooves in the bearing surface to supply the correct amount of lubricant over the surface of the journal. They are lubricated five to ten times as often as rolling bearings, using flushing action to keep out contaminants. Maintaining a bead of grease at the seals is the usual method of assuring adequate greasing intervals. Oil cups serve the same purpose in some machine designs.

21.14.3 Bearing Lubricants All classifications and formulations of oils and greases are recommended, taking advantages of specific properties or reflecting factory and field experience with available products. The high-torque and shock

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loading nature of this machinery, however, in general demands lubricants of high loadcarrying capacity and capable of operation at temperature extremes and in contaminating atmospheres. To lubricate all bearings on a farm, and particularly on a large construction site, requires knowledge of the general formulations of the lubricants and their physical characteristics. Since equipment is increasingly being lubricated through centralized systems, the pumpability of the lubricant should be questioned; in the case of greases, the ability to show a minimum of ingredient separation is paramount.

21.15 Pneumatic Equipment Compressed air is used extensively in heavy equipment to apply brakes or shift transmissions, operate gates and valves, apply clutches, operate air tools, and rotate rope drums, to mention just a few uses. Air power is as useful in many cases as hydraulic power to control and power implements.

21.15.1 Air Compressors On large farms and on construction sites, the most noticeable machines are mobile, self-contained air plants with the power source and compressor under one protective housing. They employ gasoline or diesel engines as the prime mover, generally connected to the compressor through a flexible coupling. In many designs, the compressor and engine share the same crankcase and oil pan, half of the cylinders being engine and half being compressor. Compressors in the work train of, say, cranes or concrete batching plants are smaller units, and may be belt driven off the main engine or by an electric motor. All compressors are of the positive displacement type, either reciprocating piston, rotary screw, or rotary sliding vane, or are usually limited to a maximum pressure output of 860 kPa (125 psig), with outputs ranging from 0.017 m3 /sec (35 cfm) to 0.85 m3 /sec (1800 cfm). However, there are uses for compressors with pressure outputs up to 70,000 kPa (10,000 psig) and higher. Problems with air compressors are usually due to two maintenance faults: too long a period between oil changes and ineffective air filtration. Because of the heat of compression (and heat of combustion in the case of integral engine compressors) and contaminants from intake air, manufacturer oil change recommendations should be strictly adhered to, with special consideration to high ambient temperatures and dusty atmospheres possibly calling for shorter periods for oil changes and filter maintenance. Because of the variety of designs and operating conditions, a variety of oils are recommended. Generally, engine oils of SAE 20 or 30 will lubricate satisfactorily in most cases, but particular conditions may call for ashless additive turbine type oils or even naphthenic oils of the soft, fluffy carbon deposit nature. Automatic transmission fluids are alternate recommendations in many cases, as are synthetic fluids.

21.15.2 Air Cylinders These units are lubricated and protected from corrosion by oil fed into the air stream by oilers. Lubricants are the lighter viscosity grades of inhibited hydraulic oils, engine oils, or the specially formulated pneumatic tool lubricants.

21.15.3 Air Motors As with hydraulic motors, these are usually of the vane or piston type and can do practically the same jobs and with less maintenance. They receive lubrication by airborne oil mist (vane type) or by built-in splash oilers (piston type). Most air motors are geared or work through gear reduction boxes to reduce their high revolutions per minute to the more usable speeds. Lubricants used are generally the same types as those for compressors. When the gear head is an integral part of the motor, the same oil is used, but many motors operate through separate gear boxes which may call for gear oils.

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21.15.4 Pneumatic Tools Whether hand-held or mounted on a rig, these are complex machines in the variety of machine components used to provide feed, reciprocating motion, and rotation to the working piece. Pneumatic tools employ air power in several ways. Air cylinders are used to feed the tool to the work, as in feed leg drills and stoppers. Air motors, operating through gear trains, provide rotation; through a screw arrangement or gearbox and chain, feed pressure. These applications normally call for lubricants as discussed before. It is the use of percussion force in many drills; however, that demands special lubricants. In these tools, rapid reciprocating motion of the air-driven piston hammer, the splines on the hammer and rifle bar, and the sliding drill rod shanks demand lubricants containing extreme pressure additives. The wide range of operating temperatures encountered and the presence of water, either from moist air or the water flush, call for high oxidation resistance and emulsifying (and demulsifying) capabilities of the oil. A full line of “rock drill oils” for ambient temperatures of −37 to over 43◦ C (−35 to over 110◦ F) might require oils of ISO VG 32 to ISO VG 100. Lubrication is provided by hand oiling, oil reservoirs, airline oilers, or central systems.

21.16 Good Lubrication Practice It falls on farm or construction management, the machine operator, and the fuel and lubricant supplier to see that the best lubrication practice is performed to keep complex equipment operating with a minimum of downtime, and working efficiently and safely.

21.16.1 Management Each work site must be provided with the proper number and sizes of machines and tools to do the job in the time allotted and trained operators selected. Maintenance must be accounted for in the plan, and this requires effective scheduling of fuel, lubricant, filter, and parts inventories to fit the mix of equipment on the job. Proper recordkeeping forms should be provided to control scheduled maintenance and monitor the performance and life of machine components. Lubrication schedules must incorporate the myriad oil and grease specifications and the differing relubrication intervals brought on by the large variety of machine components and configurations. Such a program may of necessity contain compromises.

21.16.2 Operators Trained not only in working his machine to its fullest utility, the operator should also thoroughly know his equipment so that he can spot slight malfunctions and recognize the need for maintenance, adjustment, or repair. Even though he may not do the actual servicing, a good operator will check all fluid levels; drain the water and dirt from oil and fuel filters and strainers; check the condition of air, oil, and water filters; and inspect compartments and lines for leaks before starting the day’s work. During operation, machine gauges and meters and the air restriction indicator should be checked frequently and engine exhaust should be monitored for color. Any highly unusual situations should be reported to the maintenance crew immediately.

21.16.3 Maintenance Preventive maintenance is receiving increasing attention for tighter cost and downtime control and to keep older machines in good shape longer. Lubrication personnel should be thoroughly familiar with the machinery drive trains and armed with the proper charts, should miss no lubrication points or filter maintenance. The operators’ reports should be thoroughly investigated. Many malfunctions may be traceable to improper lubricant or lubrication, lack of maintenance of filters, strainers and breathers, or contamination of a lubricant or fuel.

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Diagnosis of present or potential problems can often be made by examination of filters, magnetic drain plugs, engine exhaust, used lubricants, and the sound of the various components in operation. Engine oil filters may show faulty combustion and wear particles, and to the experienced person, a glycol leak. Other component filters can show excessive oxidation and contamination and wear particles that may be circulating. Diesel fuel filters may point out the presence of water in the fuel storage and the ever-present microorganisms whose waste leaves slime and sludge. Ferrous wear is readily detected by the magnetic drain plugs used in almost all components. A blue casted engine exhaust reveals excessive lubricating oil consumption; a black exhaust shows faulty combustion caused by too rich fuel–air mixtures. In the case of diesel engines, black exhaust can be caused by restricted air intake or faulty injectors. A white exhaust signals water vapor, but in diesel engines may signify raw unburned fuel or a cold smoke condition with low cetane fuel. Used lubricants, of course, readily show by their condition, color, or smell the effects of contamination and high temperature. Many test kits are available for on-the-site analysis of used oils and their judicious use may prevent some future failures. As discussed later, lubricant laboratories are well equipped with the instrumentation and technicians necessary for thorough investigation and analysis of test results. Gear or bearing whine may be a sign of wear; clunking of gears shows the presence of broken teeth. Engine sound changes may indicate fuel or air system problems. Unusual sounds in any compartment can often be traced to improper lubrication practice, either the wrong lubricant or too long a drain or greasing interval for the operating conditions, or faulty filter and strainer service. The maintenance crew, then, should not only be capable of making repairs, but should also be able to diagnose symptoms and correct the problems before catastrophic failures occur.

21.16.4 Supplier Whenever new equipment is put into operation or a new project is started on the farm or building site, the petroleum supplier should be a part of the start-up. The supplier should know the equipment, be able to follow its drivelines through, be able to interpret the lubricant designations of the machine manufacturer, and work with management and maintenance in setting up a consolidated lubricant and lubrication program for the machinery on that job. If the machines of many different manufacturers are represented on a construction site or large farm, there will possibly be a large variety of lubricants recommended (Table 21.2) because different machine makers may call for different types of lubricants for the same application (Table 21.3). Also, the intervals for the services listed in Table 21.4 will vary from one machine maker to another. Any compromises in lubricants or application intervals in the consolidated program deserve the expertise of a fuel and lubricants supplier who knows his products’ physical and chemical properties. What quality of fuel is available; what fuel system or fuel storage additives may be needed? Which multipurpose oils may be used in more than one engine make, more than one transmission make; which multipurpose greases may be used for more than one bearing application? These are just a few questions for which the supplier must have the answers.

21.16.5 Oil Analysis While the appearance and odor of used oils and filtering media may give some qualitative indication to experienced personnel of the operation of engines and other compartments, the use of field kits for more sophisticated analysis is widespread. For over four decades, methods and portable equipment have been available for field testing for water, glycol, and fuel contamination, nature of solid contaminants, and viscosity of the sample. Everyone is familiar with the hot plate test for water, the “blotter” test for solids, dispersancy and oxidation; chemical tests for glycol, microscopes for closer identification; centrifuges for solids; and various equipment for viscometry. While field tests give qualitative answers to general questions, more accurate determinations, quantitative answers and chemical analyses require well-equipped and staffed laboratories. Indeed, the past decade has seen the laboratory become a valuable consulting service

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TABLE 21.2

Farm and Construction Equipment Lubricantsa

Symbol EO

Type Engine Crankcase Oil (diesel and gasoline), described in SAE J183, J300, and J304 SA CA SB CB SC a CC a SD a CD

ATF

HTF

BF

HYDO

Automatic Transmission Fluid, described in SAE J311 Type A Type A, Suffix A Dexron® Type F Hydraulic Transmission Fluid Type C-1 a Type C-2 Brake Fluid, described in SAE J1702 and SAE J1703 SAE J1702 a SAE J1703 (formerly SAE 70R3) Hydraulic Oil MIL-H-5606 Industrial hydraulic oil resistant to rust, oxidation, and foaming industrial hydraulic oil with antiwear additives, resistant, resistant to rust, oxidation, and foaming

Symbol

Type

FRF

Fire Resistant Fluid (hydraulic) Oil/Water Emulsion Water Glycol Fluid Phosphate Ester Type Fluid Regular Type Gear Lubricant, described in SAE J306, SAE J308, and ASTM RR25-D2 (addendum 10/68) Straight Mineral Oil or API Service GL-1 Multipurpose Type Gear Lubricant, described in SAE J306, SAE J308, and ASTM RR25-D2 (addendum 10/68) API Service GL-4 or MIL-L-2105 a API Service GL-5 or MIL-L-2105B API Service GL-6 Open Gear Lubricant Track Roller Lubricanta

RGL

MPL

OGL TRL MPG MPGM WBG HTG SPC

Multipurpose Type Grease, described in SAE J310 Multipurpose Type Grease with Molybdenum Disulfide Wheel Bearing Grease, described in SAE J310 a High Temperature Grease Special Lubricant

a The specifications, classifications, or lubricants marked with an asterisk are found in common use today. It is strongly

recommended that on any single machine a minimum number of lubricants be used. It is further recommended that engine oil, multipurpose type grease, and multipurpose type gear lubricant be used wherever possible. These lubricants may be known by specific trade names or performance specifications. Military standards, MIL-H-2105, MIL-H-5606, MIL-H-21058, are available from U.S. Government, DODSSP, Standardization Documents Order Desk, Building 4D, 700 Robbins Avenue, Philadelphia, PA 19111-5094. Source: Extracted from SAE Recommended Practice J754a, in SAE Handbook, Part 2, Society of Automotive Engineers, Warrendale, PA, 2004, 40.88. With permission.

for the contractor (and possibly the large farmer), whether the consultant is the equipment dealer, the petroleum supplier, or an outside firm specializing in the service. Routine programs of analysis serve three purposes: 1. Detection of the onset and the type of abnormal wear 2. Detection of the rate and type of contamination 3. Determination of practical or reduced drain intervals Such programs call for routine sampling and testing because in all three purposes “rate” is inherent. Since no two pieces of equipment or two units are exactly the same, all machines should be sampled. This is ideal, of course, and some compromises may have to be effected by time and economics. Possibly one hydraulic system, one engine, or one gearcase may have to serve as the basis for all others. This spot approach will indicate if routine maintenance procedures are satisfactory. It may not, however, detect the machine or component in the fleet with an abnormal condition. Samples should be taken from the various systems using, at least initially, the manufacturers’ recommended drain intervals: 100 to 250 h for engines and 250 to 2000 h for hydraulic systems and other components, since these ranges cover the average change interval recommendations. Professional laboratories are better equipped for accurate qualitative and quantitative determinations; wear metals must be determined

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Lubrication of Construction Equipment Components

Component

Lubricants used

Engine crankcase (diesel and gasoline) Diesel fuel injection pump housing Air cleaner, oil bath Clutches and brakes (wet) Hydraulic wheel brake systems Hydraulic control systems Hydraulic transmissions Transmissions Bevel gear and final drive gears Limited slip differentials Gear components (other than above) Open gears Wheel bearings Bearings, shafts, levers, drivelines Track rollers Alternators, generator, electric motor

EO EO EO EO, ATF BF, EO EO, ATF, HTF, HYDO, FRF EO, ATF, HTF EO, RGL, MPL EO, RGL, MPL MPL EO, RGL, MPL, MPG MPL, OGL MPG, MPL, WGB MPG, MPGM, EO EO, TRL, MPL, MPG EO, MPG, HTG

Source: Extracted from SAE Recommended Practice J754a, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA, 2004, 40.88. With permission. 1. Several lubricants may be shown. They should not be mixed. 2. To minimize the number of lubricants used, specify engine oil, multipurpose type grease, and multipurpose type gear lubricant wherever possible. 3. Special lubricants may be required in any of the mentioned components. 4. For specific recommendations consult equipment manufacturer. 5. Maintenance intervals are listed in Table 21.4.

TABLE 21.4

Maintenance Intervals

Interval time in hours 10 50 100 250 500 1000 2000

Equivalent time Each shift Weekly 2 weeks Monthly 3 months 6 months Yearly or whichever occurs first

Note: It is recommended to consult manufacturer and/or the operations & maintenance manual. Source: Extracted from SAE Recommended Practice J753, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA, 2004, 40.78. With permission.

in them because of the sophisticated spectrographic equipment needed. With data on the metallurgical design of the various engines and other systems, laboratories can more readily detect sources of wear and contamination, analyze the rates of generation, and provide consultation on maintenance and repairs needed to prevent catastrophic failures. “Catastrophic” is a widely used term or adjective to denote sudden, violent, widespread damage as opposed to normal wear over a period of time. Oil analysis, consultation on maintenance, and repairs are to prevent catastrophies. Best lubrication practice, then, keeps machines running with a minimum of downtime for maintenance and repairs, for the longest possible life. Farm and construction management, their maintenance

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personnel, and the petroleum supplier all have vital roles in this effort, and they must be cooperative if the production intended for man’s machines is to be reached and improved.

Acknowledgment We acknowledge the initial work by William J. Hanley in the previous volume.

References [1] SAE Recommended Practice J183 Engine Oil Performance and Engine Service Classification (other than "energy conserving"), in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.2. [2] SAE Recommended Practice J300 Engine Oil Viscosity Classification, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.33. [3] SAE Recommended Practice J304 Engine Oil Tests, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.36. [4] SAE Recommended Practice J311 Fluid for Passenger Car Type Automatic Transmissions, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.62. [5] SAE Recommended Practice J1702 Self-propelled Sweepers Sweep-Ability Performance, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.23. [6] SAE Recommended Practice J1703 Motor Vehicle Brake Fluid, in SAE Handbook, Volume 2, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.54. [7] SAE Recommended Practice J306 Automotive Gear Lubricant Viscosity, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.55. [8] SAE Recommended Practice J308 Axle and Manual Transmission Lubricants, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.52. [9] SAE Recommended Practice J310 Automotive Lubricating Greases, in SAE Handbook, Volume 1, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 12.56. [10] SAE Recommended Practice J754a Lubricant Types — Construction and Industrial Machinery, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.88. [11] SAE Recommended Practice J753 Maintenance Interval Chart, in SAE Handbook, Volume 3, Society of Automotive Engineers, Warrendale, PA 15096-0001, USA, 2004, 40.78. [12] Imperial Oil, Product Information: Lubricants and Specialties, 10th ed., 1999. [13] http://www.nass.usda.gov/census/. Accessed on 14 March 2005. [14] http://www.census.gov/mcd/. Accessed on 14 March 2005.

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22 Industrial Lubrication Practice — Wheel/Rail Tribology 22.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.2 Wheel/Rail Contact Mechanics . . . . . . . . . . . . . . . . . . . . . . .

22-1 22-2

Contact Position • Friction and Creep • Contact Stress • Wheel and Rail Profiles

22.3 Wheel and Rail Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.4 Wheel and Rail Damage Mechanisms . . . . . . . . . . . . . . . .

22-7 22-8

Wear and Plastic Deformation • Wear Transitions • Wear Mapping • Rolling Contact Fatigue • Interaction of Wear and Fatigue • Modeling Damage Mechanisms

Roger Lewis and Rob Dwyer-Joyce Department of Mechanical Engineering, The University of Sheffield

22.5 Friction Modification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

22-17

Increasing Friction • Reducing Friction

22.6 Maintenance of Wheels and Rails . . . . . . . . . . . . . . . . . . . . 22.7 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

22-20 22-20 22-21

22.1 Introduction The wheel/rail contact is a crucial component in the successful operation of railways. A large variety of loading conditions and contact geometries exist due to the many different rail and wheel profiles, rail cant and curve radii, and railway vehicles running on a network. Contact conditions vary considerably between the two main contact areas: wheel tread/railhead and wheel flange/rail gauge corner, but are usually more severe in the latter, where greater wear and fatigue cracking is seen to occur. Friction and creepage in the contact are also highly variable. Natural lubricants such as humidity, precipitation, and leaves can negatively influence the friction in the wheel/rail contact, causing braking problems and wheel slip in traction. These problems can be overcome by using applied lubricants to reduce wear in curves and friction modifiers to increase adhesion. This, however, further adds to the complexity of the wheel/rail contact system. Effective management of the wheel/rail contact is an important aspect of rail infrastructure operations. Rail maintenance was estimated to have cost 300 million Euro within the European Union in 1995 (Cannon, 1996). All the influencing factors have to be taken into account as they interact closely (as indicated in Figure 22.1) (adapted from Kalousek and Magel, 1997). For example, measures used

22-1

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22-2

Handbook of Lubrication and Tribology

Rail and wheel life up

Wear

Plastic flow

Damage modes Rail rollover

RCF

Hollow wheels

Profit up

Friction management

Thermal cracks

Wheelset dynamics

Maintenance costs down

Wheel/rail material

Contact mechanics Spending down

FIGURE 22.1 Systems approach to wheel/rail interface management and research. (Adapted from Kalousek, J. and Magel, E., 1997, Railway and Track Structure, January.)

to reduce wear, such as lubrication, can influence fatigue and adhesion, and the measures used to increase adhesion, such as sanding, can have a detrimental effect on wear. A fine balance has to be found in determining maintenance schedules and lubrication regimes to keep railway networks running smoothly. This is becoming increasingly difficult as new specifications on wear and reliability are being imposed to increase the time between reprofiling, increase safety, and reduce total life-cycle costs. In parallel with these requirements, vehicle missions are changing due to the need to operate rolling stock on track with low radius curves, as well as the high radius curves found on high speed lines and increasing speeds. These are leading to an increase in the severity of the wheel/rail contact conditions (Stanca et al., 2001).

22.2 Wheel/Rail Contact Mechanics 22.2.1 Contact Position The position of the wheel/rail contact, which is typically 1 cm2 in size, varies continuously as a train progresses down a section of track. The exact position will depend on the wheel and rail profiles and the degree of curvature of the track and whether the wheel is the leading or trailing wheelset on a bogie, as well as other factors determined by the bogie design. In straight track, however, it is likely the wheel tread and railhead will be in contact with wheel flange and rail gauge corner contact occurring in curved track. Figure 22.2 shows how the contact position and stress varies for the two wheels on a leading wheelset entering a right-hand curve. Three possible regions of wheel/rail contact have been defined (as shown in Figure 22.3) (Tournay, 2001): 1. Region A — Wheel tread/railhead. The wheel/rail contact is made most often in this region and usually occurs as the railway vehicle is running on straight track or very high radius curves. This region yields the lowest contact stresses and lateral forces. 2. Region B — Wheel flange/rail gauge corner. The contact in this region is much smaller than that in region A and is often much more severe. Typically contact stresses and wear rates are much higher. If high wear and material flow occurs, two point contacts may evolve, where tread and flange contact is apparent.

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Industrial Lubrication Practice

22-3

Contact stress

FIGURE 22.2

Leading wheelset entering a right-hand curve.

Region C

Region A

Region B

FIGURE 22.3 Wheel/rail contact zones. (Adapted from Tournay, H., 2001, Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.)

3. Region C — Contact between field sides of wheel and rail. Contact is least likely to occur here and if it does, high contact stresses are induced, which will lead to undesirable wear features causing incorrect steering of the wheelset.

22.2.2 Friction and Creep The friction between the wheels and rail is extremely important as it plays a major role in the wheel/rail interface process such as adhesion, wear, rolling contact fatigue, and noise generation. Effective control of friction is managed through the application of friction modifiers to the wheel/rail contact. The aim of friction management is to maintain friction levels in the wheel/rail contact to give: low friction in the wheel flange/rail gauge corner contact; intermediate friction wheel tread/railhead contact (especially for freight trucks); and high friction at the wheel tread/rail top contact for locomotives (especially where adhesion loss problems occur) (Zakharov, 2001). Ideal friction conditions in these contact regions for high and low rails are shown in Figure 22.4 (Roney, 2001; Sinclair, 2004). The wheel/rail contact occurs in the incipient sliding regime, hence there is not complete adhesion within the contact, and stick and slip zones are apparent, as shown in Figure 22.5. Three types of slip (or creepage) occur in the wheel/rail contact; lateral (perpendicular to the direction of wheel motion), longitudinal (in the direction of wheel motion), and spin creepages and these are generated as the wheel deviates from a pure rolling motion. These vary in the same way as contact stresses change as the wheel/rail contact position moves and they are more severe (particularly lateral creepage) in curves. The degree of creepage in the contact depends on the normal load and friction in the contact, as shown in Figure 22.5, which illustrates a creep curve for a wheel/rail contact. The greater the traction force, the larger the slip region in the contact. This curve varies dramatically with the introduction of a third body layer to the contact, such as a lubricant (natural or applied) or a friction modifier.

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22-4

Handbook of Lubrication and Tribology 0.25 < m < 0.4 0.25 < m < 0.4 m < 0.1

Low rail

High rail

FIGURE 22.4 Ideal friction coefficients in the wheel/rail contact. (Adapted from Sinclair, J., 2004, Vehicle Track Interaction: Identifying and Implementing Solutions, IMechE Seminar, February 17th.)

Creep = 0.01 to 0.02

Rolling direction

Tractive force (= mN)

Slip

Tractive forces Slip

Stick

Stick Stick

Slip

Slip Creep

FIGURE 22.5

Relationship between traction and creep in the wheel/rail contact.

22.2.3 Contact Stress As a result of the fact that the contact position is not spread evenly over the entire wheel or rail profile, the shape of the profiles will change as time progresses, due to wear and material flow (processes largely controlled by the loads and creepage within the contact). In order to be able to predict how profiles may evolve, a good understanding of the contact stress is therefore required. The simplest solution for determining wheel/rail contact geometry and stress is Hertz analysis (see Johnson [1985]), where the wheel and rail can be equated to two cylinders in contact perpendicular to each other. The maximum contact pressure, p, is given by:
p=

3

3PE 2 2π 3 R 2 (1 − ν 2 )2

(22.1)

where P is the normal load, E and ν are the Young’s modulus and Poisson’s ratios respectively (assumed to be same for wheel and rail materials in this case), and R is the equivalent radius given by: 1 1 1 = + R R1 R2

(22.2)

where R1 and R2 are the contact radii of the wheel and rail. This approach is, however, limited in accuracy due to the assumptions made in the analysis, such as smooth contacting surfaces, a linear elastic material response and that the contact dimensions must be small compared to the radii of curvature of

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Industrial Lubrication Practice

22-5

FEM y

CONTACT

Hertz

(Case 1, N = 80377 N) x

PMAX = 1500 MPa

PMAX = 3457 MPa

Dx(max) = 20 mm Dy(max) =17 mm

Area = 134.8 mm2

PMAX = 3601 MPa

Dx(max) =18.1 mm Dy(max) =11.3146 mm

Area(sum) = 46.7181 mm2

Dx = 19.6 mm Dy = 12 mm

Area = 45 mm2

(Case 2, N = 80377 N) PMAX = 665 MPa Area = 172.8 mm2

PMAX = 865 MPa Area = 134.21 mm2

PMAX = 1080.53 MPa Area = 111.58 mm2 x y

Dx = 12.53 mm Dy = 18.58 mm

Dx = 10.12 mm Dy = 15.45 mm

Dx = 10.34 mm Dy = 13.74 mm

FIGURE 22.6 Comparison between FE, CONTACT, and Hertz analysis. (Adapted from Telliskivi T. and Olofsson U., 2001, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 215, pp. 65–72.)

the contacting bodies. In the flange contact particularly, the contact radius can be as small as 10 mm, which means this assumption can be invalid. Numerical solvers have been developed for calculating area and stress for contacts approximated to Hertzian ellipses, such as FASTSIM (Kalker, 1982) and non-Hertzian contacts, such as CONTACT (Kalker, 1990). CONTACT, however, requires a large amount of computing resource and again is limited by the half-space assumption. Finite element modeling carried out by Telliskivi and Olofsson (2001), including the plastic deformation and actual wheel and rail profiles showed good correlation in terms of contact area and stress with Hertz and CONTACT for the railhead contact, but considerably different stress and area for the rail gauge corner flange contact (as shown in Figure 22.6, where Case 1 represents the gauge corner contact and Case 2 represents the head contact), due to the limiting half-space assumption in Hertz and CONTACT analysis. Recently an innovative ultrasonic technique has been used to study the wheel/rail contact (Marshall et al., 2004). Figure 22.7 shows contact pressures derived from ultrasonic scans compared with numerical calculations using actual roughness profiles and Hertz analysis. There is good global geometric correlation between the ultrasonic results and the numerical model. The degree and fragmentation of the ultrasonic and numerical contacts are qualitatively similar. But on a local

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22-6

Handbook of Lubrication and Tribology

(a)

(b)

2200 2000 1800

1400 1200 (c)

(d)

1000 800

Contact pressure (MPa)

1600

600 400 200 0

FIGURE 22.7 Contact pressure maps for a load of: 80 kN (a) ultrasonic measurement; (b) Hertzian; (c) elastic model; (d) elastic–plastic model. (Adapted from Marshall, M.B., Lewis, R., and Dwyer-Joyce, R.S., 2004, Ultrasonic characterisation of a wheel/rail contact, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Triboloby Series No. 43, pp. 151–158.)

level the ultrasonic results and numerical solutions differ. This is likely due to difficulty in aligning the surfaces to the same orientation in both the experiment and model. The elastic case assumes no localized yielding at the contact; this results in predicted contact pressures in excess of yield for the contacting surfaces. The pressures determined for the elastic case are far in excess of those measured ultrasonically. However, the elastic–plastic case shows similar peak pressures to the experiment. The experimental and elastic–plastic numerical model peak pressures are in excess of the Hertz solution, this is due to the reduced contact conformity attributable to roughening. Hertzian theory dictates that in static loading the maximum compressive stress is at the surface and the maximum shear stress is below the surface (at a depth of 0.78a, where a is the contact half width). When tractive force is applied at the surface the shear stress increases and the position of the maximum value moves closer to the surface. Because of the rolling/sliding behavior of a wheel on a rail, a cyclic build-up of plastic deformation occurs beneath the material surfaces. It is this behavior that leads to rolling contact fatigue and wear occurring. Figure 22.8 shows a shakedown map, which illustrates the relationship between friction in the wheel/rail contact and the load carrying capacity of the contact. It shows the limits of material behavior in terms of nondimensional contact pressure, p0 /k as a function of friction coefficient, µ (=T /N ), where p0 is the normal contact pressure, k is the shear yield strength, T is the tractive force, and N is the normal load. At relatively low friction coefficients, cumulative plastic flow occurs subsurface. For friction coefficients above about 0.3, plastic flow is greatest on the surface. The worst position in terms of damage to the material is in the ratchetting region, where strain is accumulated until the ductility of the material is exceeded and it is lost as wear debris or a crack is initiated.

22.2.4 Wheel and Rail Profiles Wheel and rail profiles are designed to optimize their performance for their given application. Performance is generally assessed in terms of (Zakharov, 2001): resistance to damage mechanisms (wear, fatigue, etc.), minimization of noise, and maximization of vehicle stability.

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Industrial Lubrication Practice

22-7

6 Alternating plasticity (plastic shakedown)

5

Load factor, p0 / k

4

Incremental growth (ratcheting)

Elastic shakedown 3

2 Elastic 1 Subsurface

Surface

0.2

0.4

0.6

Friction coefficient, m

FIGURE 22.8

Shakedown map. TABLE 22.1 Chemical Composition (weight %) of Rail Steels Elements C Mn Si S P Cr V Ni Mo

USA, Canada, Brazil

Australia

Europe (UIC60)

0.72–0.82 0.80–1.10 0.10–0.60 0.037 max. 0.035 max. 0.25–0.50 0.03 max. 0.25 max. 0.10 max.

0.72–0.82 0.80–1.25 0.15–0.58 0.025 max. 0.025 max. — — — —

0.60–0.82 0.80–1.30 0.30–0.90 0.025 max. 0.025 max. 0.80–1.30 — — —

Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.

In terms of contact stress, this means that stresses need to be kept low and contact points need to be spread across the wheel tread and railhead to reduce concentrated wear occurring. This, however, is difficult to achieve. A piece of track will see many different wheel profiles in a given length of time, which will result in many different contact positions and stresses. The main aim though is to design the profiles such that regrinding intervals are as long as possible, to reduce costs.

22.3 Wheel and Rail Materials Rail and wheel steels are metallurgically very similar. Both utilize high carbon (0.65 to 0.82%) and have a pearlitic microstructure. Some typical chemical compositions of rail steels from around the world are shown in Table 22.1. There are many different specifications of wheel steels, the main difference between them being carbon content. Choice varies according to final use, which may be freight, passenger, or locomotive. Some typical examples are shown in Table 22.2.

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22-8

Handbook of Lubrication and Tribology TABLE 22.2 Steels Elements C Mn Si S P Cu V Ni+Cr+Mo

Chemical Composition (weight %) of Wheel

USA, Canada, Brazil (Class C)

Australia

Europe (R7T)

0.67–0.77 0.60–0.85 0.15 max. 0.05 max. 0.05 max. — — —

0.67–0.77 0.60–1.00 0.15 max. 0.035 max. 0.04 max. — — —

0.51 0.77 0.35 0.009 0.009 0.14 0.001 0.20

Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.

TABLE 22.3

Mechanical Properties of High Strength Rail Steels

Property Yield strength (MPa) (min.) Tensile strength (MPa) (min.) Elongation % (min.) Brinell surface hardness

USA, Canada, Brazil

Europe (UIC60)

758 1172 10 340–390

640 1080 9 320–360

Source: Adapted from Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA.

The mechanical properties of rail steels of most interest are: yield strength, which dictates the plastic flow and work hardening characteristics of the material; tensile and fatigue strength, which is an indication of the material’s resistance to fatigue and hardness, which can indicate how the material will resist wear. Typical rail mechanical properties are shown in Table 22.3. While the majority of rail steel is currently pearlitic, bainitic rail steels have been developed to try and improve damage resistance of track (Kalousek et al., 1985; Ueda et al., 1997; Singh et al., 2001). Whereas these offer greater wear resistance than pearlitic steels, their resistance to rolling contact fatigue is lower (Jiang, 1999; Yokoyama, 2002; Sawley, 2003). Work is, however, continuing on their development with a view to overcoming this problem.

22.4 Wheel and Rail Damage Mechanisms A number of damage mechanisms exist for both rail and wheels. The most significant are wear, plastic deformation, and rolling contact fatigue (RCF). These individually can cause problems, but due to their close interaction, measures introduced to reduce one may increase another. In this section, each mechanism is looked at in detail and then the interaction of wear and RCF is outlined.

22.4.1 Wear and Plastic Deformation A number of different techniques have been used for studying wear of railway wheel and rail steels. Field measurements have been used in the past to study the causes of wheel and rail wear (Dearden, 1960). A large amount of data has also been gathered from simulated field experiments carried out on specially built test tracks (Steele, 1982). Laboratory methods used range from full-scale laboratory experiments

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Industrial Lubrication Practice

22-9

(McEwen and Harvey, 1985) and scaled-down tests (Kumar, 1984) to bench tests using a twin disc setup (Beagley, 1976; Bolton et al., 1982; Bolton and Clayton, 1984; Krause, 1986; Garnham, 1992; and DwyerJoyce, 2004a). The twin disc approach has been used more than most because it offers greater control over experimental variables as well as the ability to test a wide range of materials at lower cost. Twin disc testing carried out to study the wear behavior of railway wheel and rail steels has led to the identification of a number of wear regimes (Beagley, 1976; Bolton et al., 1982; Kumar et al., 1984; Bolton and Clayton, 1984; Lewis and Dwyer-Joyce, 2004a). Early tests demonstrated that two wear regimes existed (Beagley, 1976; Bolton et al., 1982). These were designated mild and severe. Later work led to the identification of a further regime, designated catastrophic wear (Bolton and Clayton, 1984; Lewis et al., 2004). Figure 22.9 shows results of twin disc testing of R8T wheel steel against UIC60 900A rail steel (data from Lewis and Dwyer-Joyce, 2004a). The results are plotted in terms of wear rate (µg material lost/ m rolled/mm2 contact area) against an index based on work done in the contact, T γ /A, where T is tractive force (normal load divided by friction coefficient), γ is slip, and A is contact area. This is a useful way to plot the data as it allows comparison of different test geometries. The curve is typical of those for rail steels (Lewis and Olofsson, 2004), as shown in Figure 22.10. As can be seen wear rates are gradually reducing with time as new and more wear resistant rail materials have been introduced. Wheel tread and railhead wear are thought to fall within the mild regime and wheel flange and rail gauge corner wear in the severe or catastrophic regime. This has been verified with field measurements carried out by Olofsson and Nilsson (2002). As a result of twin disc tests (Lewis and Dwyer-Joyce, 2004a), the regimes have been characterized in terms of wear rate and wear mechanism. At low T γ /A, in the mild wear regime, oxidative wear occurs on both wheel and rail discs. The disc surfaces turned a rusty brown color. Closer examination of the wear surfaces revealed abrasive score marks and evidence of the oxide layer breaking away from the surface (see Figure 22.11[a]). This ties in with observations made in the field that on straight track where low slip occurs on the high rail, oxidative wear is prevalent generating magnetite (Fe3 O4 ) (Broster et al., 1974) and in full-scale test-rig results, where reddish oxide film appeared for low slip conditions (McEwen and Harvey, 1985). As T γ /A was increased, the wear mechanism of the wheel discs altered. The wheel disc appeared to be wearing by a ratcheting process (deformation followed by crack growth and subsequent material removal). Closer examination of the wheel disc surfaces revealed that this was the case (see Figure 22.11[b]).

3,000

Wear rate (mg / m / mm2)

2,500 2,000 Catastrophic

1,500 1,000 Mild

Severe 500 0 0

20

40 Tg /A

60

80

100

120

(N / mm2)

FIGURE 22.9 Wear rates and regimes for R8T wheel steel tested against UIC60 900A rail steel. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.)

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Handbook of Lubrication and Tribology

Wear rate (mg / m / mm2)

100,000

BS11 vs. Class D (Bolton and Clayton, 1984)

10,000 1,000

Standard carbon rail (Danks, 1987)

100 10

UIC60 900A vs R8T (Lewis and Olofsson, 2004)

1 0.1 0.01 0

50

100

150

200

Tg /A (N / mm2)

FIGURE 22.10

Rail steel wear rates.

Region of delamination of oxide film (a)

Abrasive score marks (b)

FIGURE 22.11 Wheel disc surface run at (a) T γ /A = 0.21 and (b) T γ /A = 4.1. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J, Journal of Engineering Tribology, Vol. 218, pp. 467–478.)

Figure 22.12[a] shows a section through a wheel disc, run at low T γ /A, parallel to the rolling direction. At the surface the oxide layer is just visible. There is a very small amount of deformation just below the wear surface of the disc. At higher levels of T γ /A, observation of the subsurface morphologies revealed that a larger amount of plastic deformation was occurring below the wheel disc wear surface (see Figure 22.12[b]) and crack formation just below the surface was visible, which was leading to thin slivers of material breaking away from the surface. The slivers have a similar thickness to the oxide layer and could indicate a severe oxidative wear mechanism occurring, where fracture occurs between the oxide layer and the metal. As T γ /A was increased further far greater cracking was visible at the wear surface and some of these cracks were seen to alter direction from running parallel to the wear surface and turning up to turning down into the material causing larger chunks of material to break away (see Figure 22.12[c]).

22.4.2 Wear Transitions While rail steel wear regimes have been defined well in terms of wear rate, metallographic features, and wear debris, there was not a great understanding of what mechanisms are leading to the changes in wear rate that occur (see Figure 22.9). In order to further understand the wear mechanisms, the transitions

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Industrial Lubrication Practice

22-11

(a) Oxide layer

Extent of subsurface deformation

20 mm

Rolling direction

(b)

40 mm Crack formation (c)

40 mm

More cracking and larger chunks of material breaking away

FIGURE 22.12 Sections parallel to rolling direction through wheel disc run at (a) T γ /A = 0.21; (b) T γ /A = 16.6; and (c) T γ /A = 28.3. (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.)

between these regimes were studied in more detail (Lewis and Dwyer-Joyce, 2004a). It was proposed that the first transition is associated with the onset of fully sliding contact conditions and the second is a result of surface temperature effects. Figure 22.13 shows friction measurements taken during UIC60 900A vs. R8T wear tests carried out at 1500 MPa plotted against slip, that is, a creep curve. As would be expected the friction reaches a threshold. This transition represents the change from partial slip in the disc interface to full slip conditions. Also shown is the Carter creep curve for an assumed limiting friction of 0.5. This model creep curve is based on smooth elastic cylinders in contact (Carter, 1926). The wear data (also shown in Figure 22.13) follows a similar pattern indicating that at the point of transition from partial slip to full slip a wear transition also occurs. After the full slip condition has been reached, increasing the magnitude of slip has no effect. Calculations were carried out to determine temperatures at a twin disc contact for the UIC60 900A rail steel vs. R8T wheel steel (Lewis and Dwyer-Joyce, 2004a; Lewis and Olofsson, 2004). The results, shown in Figure 22.14, indicate that the transition from severe to catastrophic wear occurs around 200◦ C. These temperatures correspond with those causing a drop in the yield strength of carbon manganese steels similar to rail steels (British Steel Makers Creep Committee, 1973).

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22-12

Handbook of Lubrication and Tribology 10,000

1

Wear rate (mg / m / mm2)

Carter creep curve 0.8

1,000

0.6 m

100 0.4 10

Wear rate

0.2

Measured m 1

0 0

5

10

15

20

Slip (%)

FIGURE 22.13 Friction vs. slip in the twin disc contact (for tests carried out at 1500 MPa). (Adapted from Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J: Journal of Engineering Tribology, Vol. 218, pp. 467–478.) 400

3500 3000

300

Total max. temp.

2500

250 2000 200 1500

150 Total avg. temp.

100 50

1000 Wear transition

Wear rate

0 0

20

40

60

80

100

Wear rate (mg/m/mm2)

Temperature (8C)

350

500 0 120

Tg/A (N/mm2)

FIGURE 22.14 material.

Twin disc contact temperatures and wear coefficients for UIC60 900A rail material vs. R8T wheel

22.4.3 Wear Mapping Although using the T γ /A method for plotting wear rate data enables wear transitions to be identified easily and comparisons of different material combinations to be made, it does not help in fully understanding how the individual contributions of different parameters such as contact pressure and slip affect wear rate. Lewis and Olofsson (2004) proposed a wear mapping method that would give a more complete analysis of the effect of individual parameters. It is based on the technique developed by Lim and Ashby (1987) for mapping sliding wear mechanisms. Wear coefficients were calculated from the rail steel wear data using Archard’s equation (Archard, 1953): Vh K = (22.3) Ns where K is the wear coefficient, V is the wear volume, N is the normal load, s is the sliding distance, and h is the material hardness. Wear coefficients were then plotted against contact pressure and sliding speed in the contact. Two types of plots were constructed; contour maps and 3D point graphs (see Figure 22.15). Obviously, the accuracy of the contour map is limited by the amount of data available. The accompanying 3D graphs

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Industrial Lubrication Practice Severe

40

1000 900 800

hic transition

1100

)

–4

Catastrophic

Severe – catast rop

Contact pressure (MPa)

1300 1200

35

30.63 -- 35.00 26.25 -- 30.63 21.88 -- 26.25 17.50 -- 21.88 13.13 -- 17.50 8.750 -- 13.13 4.375 -- 8.750 0 -- 4.375

Wear coefficient (×10

Mild 1400

22-13

30 25 20 15 10 5

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

0.8 0.6 1200

Con

tac

700 0.0

1.0

0 1400 0.4 1000

t pr

ess

ure

800 (MP a)

0.2 0.0 600

Sliding velocity (m/sec)

c)

se

m/

y(

it loc

e gv

din

Sli

FIGURE 22.15 Wear coefficient maps for UIC60 900A rail material vs. R7 wheel material data from Olofsson and Telliskivi (2003). (Adapted from Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729.) Railhead/wheel tread

Rail gauge / wheel flange

Contact pressure (MPa)

2500

2000

Severe — catastrophic transition

1500

UIC60 900A vs. R7 wear map

Mild Catastrophic

1000

500

Severe

0 0.0

0.2

0.4

0.6

0.8

1.0

Sliding velocity (m/sec)

FIGURE 22.16 UIC60 900A rail steel wear map plotted over wheel/rail contact conditions derived from GENSYS simulations. (Adapted from Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729.)

give an indication of where data is lacking on a particular map. Wear transitions were marked on the contour plots. To study how the wear regimes identified above fit in with the wheel/rail contact conditions shown in Figure 22.15 the wear map of UIC60 900A rail steel vs. R7 wheel steel has been overlaid, as shown in Figure 22.16 (Lewis and Olofsson, 2004). This indicates that the railhead/wheel tread contact will experience mild to severe wear and the rail gauge/wheel flange contact will experience severe to catastrophic wear. This backs up previous suppositions regarding the wear regimes into which the railhead/wheel tread and rail gauge/wheel flange contacts fall.

22.4.4 Rolling Contact Fatigue Fatigue failures for rails and wheels can be categorized into surface and subsurface failures. Surface fatigue in rails can lead to head checking or squat formation and subsurface fatigue can result in shelling or the formation of tache ovale.

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22-14

Handbook of Lubrication and Tribology

Crack initiation and propagation by ratcheting

Crack driven by bending and residual stresses

Contact stresses dominate

(da/dn)

Crack driven by ratcheting

Crack driven by contact stress field Bending stress dominates crack propagation

Crack length

FIGURE 22.17 Phases of crack growth in rail. (Adapted from Kapoor, A., Fletcher, D.I., and Franklin, F.J., 2003, Proceedings of the 29th Leeds-Lyon Symposium on Tribology, pp. 331–340.)

Motion

Fluid filled crack

FIGURE 22.18

Crack opening driven by fluid pressurization.

Head checks (shallow cracks at the rail surface) normally occur on the rail gauge corner on curves. They result from accumulation of plastic strain increments (ratcheting), which eventually exhaust the ductility of the surface material, at which point cracks can initiate. The critical contact conditions for this to occur are high load and friction (see ratcheting region in the shakedown map shown in Figure 22.8). Plotting surface crack growth rates (da/dn) against crack length reveals that there are a number of phases present as shown in Figure 22.17 (Kapoor et al., 2003). After initiation, crack growth is driven by ratcheting in the plastically deformed layer. As the crack becomes longer and deeper crack growth is driven by the stress field due to the repeated contact loading. Finally the crack turns downward and growth is driven by bending stresses in the rail. If the crack reaches a critical crack length at this stage fast fracture can occur resulting in a rail break. Water and lubricants trapped in a crack increase the speed of propagation (Bower, 1988; Bogdanski et al., 1996) as shown in Figure 22.18. This is because liquids are trapped in the rail cracks as the wheel passes over causing pressurization, which increases the crack growth rate. Work has been undertaken to study the effect of rail surface indentations (Gao, 2000) and contaminants such as ballast dust and sand applied to increase adhesion in the wheel/rail contact (Grieve et al., 2002; Lewis, 2004) on rail fatigue life. Tests were carried out under various states of lubrication. These indicated that in dry or water lubricated conditions surface indentations had negligible effect on the fatigue life and that the dents were removed by plastic flow of surrounding material. For oil lubricated conditions, however, surface damage acted as a fatigue initiation site (Gao, 2000). For the contacts contaminated with ballast dust, wear rates were enhanced. The particles were seen to embed in the softer wheel material and abrade the harder rail surface (Grieve et al., 2002). Tests with sand showed very high wear rates. In addition to abrading the surfaces the sand initiated a low cycle fatigue process, which resulted in the removal of fatigue spalls from the material surface (Lewis and Dwyer-Joyce, 2004b).

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Industrial Lubrication Practice (a)

22-15 (b) Fatigue crack Point of propagation initiation

Initial wheel surface

Fatigue crack

Final fracture

FIGURE 22.19

RCF damage: (a) surface fatigue; (b) crack morphology for a subsurface induced fatigue failure.

Squats occur on straight track on the surface of the railhead. They appear as darkened areas on the rail. They consist of two cracks, one in the direction of travel and the other in the opposite direction, which is much longer. They can initiate as a result of ratcheting and fluid pressurization and also from white etching layers (WELs). WELs result from modification of the microstructure of the rail surface material from pearlite to martensite (Pyzalla et al., 2001). They not only have been noted to have a hardness of up to 1200 Hv (Feller, 1991), but are also extremely brittle and crack initiation is therefore more likely to occur within a WEL. WELs normally occur as a result of high temperatures, which may result from wheel skids, for example. Shelling in rails occurs at the rail gauge corner in curves and is a subsurface initiated defect (Grassie, 1997). Elliptical shell-like cracks propagate parallel to the rail surface and in many cases cause material to spall away. Tache ovale are defects which develop about 10 to 15 mm below the surface of the railhead from cavities caused by the presence of hydrogen (Grassie, 1997). The cavities may be present in the parent rail material or can arise during welding if it is carried out poorly. If the crack becomes sufficiently large, transverse fracture of the rail can occur. In surface initiated fatigue of railway wheels, fatigue cracks result from excessive plastic flow of the surface material. This will cause crack initiation due to ratcheting or low cycle fatigue of the surface material on both. Once initiated, the cracks typically grow into the wheel material to a maximum depth of 5 mm. Final fracture occurs as the cracks branch toward the wheel tread. The typical appearance of surface initiated fatigue failure is shown in Figure 22.19(a). Surface initiated cracks are normally not a safety issue. However, they are the most common type of fatigue damage in wheels. They are costly in requiring reprofiling of the wheel and causing unplanned maintenance. In the case of subsurface fatigue, cracks will initiate several millimeters below the surface. They continue to grow under the surface and final fracture will normally occur due to branching toward the surface. Such a failure will lead to a large piece of the wheel surface breaking away, as shown in Figure 22.19(b).

22.4.5 Interaction of Wear and Fatigue If the effect of wear on crack propagation is considered (as shown in Figure 22.20[a]), the crack is truncated as more material is removed. If crack truncation rate is laid over the crack growth rates the net crack growth rate can be determined. In Figure 22.20(b), two different wear rates are considered. If the wear rate (and hence crack truncation rate) is high, it is likely that most cracks will be worn away before progressing beyond Stage A (cracks driven by ratcheting). If, however, the wear rate is low, cracks will initiate and progress along Stage A. They may stabilize at Point 1, where growth rate equals truncation rate. These curves will vary considerably with contact conditions and as such the crack may be carried from Point 1 to Point 2 and then continue propagating. The crack reaches a length at which the growth

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Handbook of Lubrication and Tribology (a)

Material removed by

Truncation of crack = depth of material

Crack at angle u

(b)

(c) Life line due to fatigue

Crack truncation due to high wear rate

A

C Life

(da/dn)

B

1 2

3 Crack truncation due to low wear rate Crack length

Life line due to wear Fatigue failure unsafe

Failure by wear safe

Material removal rate (by grinding or wear)

FIGURE 22.20 Interaction of wear and fatigue: (a) crack truncation by wear; (b) crack growth rate vs. crack length; (c) rail life vs. material removal rate.

rate declines and it stabilizes at Point 3. If the wear rate drops below the intersection of Stages B and C, the crack can move to Stage C, which may lead to a dangerous conclusion. Figure 22.20(c) shows the effect of material removal rate on rail life. This may be reduced by using harder rail or by lubrication. Grinding rail increases the removal rate. Clearly the aim should be to have a removal rate at the optimum point, where the wear and fatigue lines cross. This, however, will be very difficult because of variations in the wheel/rail contact conditions, but with good interface management systems in place a good life can be attained.

22.4.6 Modeling Damage Mechanisms A number of design tools have been created to predict damage mechanisms in the wheel/rail contact (Fries and Dávila, 1987; Pearce and Sherratt, 1991; Zobory, 1997; Jendel, 2000a; Lewis et al., 2004). Typically wheel profile evolution prediction tools incorporate dynamics modeling to predict global contact parameters to assess contact position, load, and slip; local contact modeling based on, for example, FASTSIM or CONTACT, to calculate slip and load distributions within the contact and a semiempirical wear model based on coefficients derived from rolling/sliding or sliding wear tests (as shown in Figure 22.21). Examples of these tools include the wheel durability predictor based on ADAMSRail multibody dynamics software, developed during the EC funded HIPERWheel project (Lewis et al., 2004) and the wheel wear assessment code developed by Jendel (2000). The wear model in the HIPERWheel work was based on the T γ approach and that used by Jendel involves wear coefficients from pin-on-disc testing. The HIPERWheel predictor also included a fatigue assessment tool called FIERCE (Fatigue Impact Evaluator for Rolling Contact Environments), created by Eckberg (2002). The three mechanisms of fatigue initiation: surface, subsurface, and deep defect, are considered separately and quantified by three fatigue indices (details in Eckberg [2002]). The fatigue indices evaluated by FIERCE are represented as histograms. In addition, the surface fatigue impact may also be represented in the form of work-points plotted in a shakedown diagram. Interaction of fatigue and wear is not, however, considered. The work initiated by Jendel, has been taken forward and wear predictions now include the effects of natural lubrication and braking (Enblom, 2004).

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Industrial Lubrication Practice

22-17

Wheel and rail profiles

Multi-body dynamics model

Local contact analysis

Wear model

Change in wheel profile

FIGURE 22.21 Wheel profile evolution prediction tool schematic.

The two main drawbacks in some of these analysis tools are in the contact mechanics method used and the approach to wear modeling. Telliskivi and Olofsson (2001, 2004) studied different contact mechanics methods used in wear simulations and concluded that they significantly affected the accuracy. They also developed a new method with increased accuracy and no loss of calculation speed. Wear models are typically semiempirical and based on the results of laboratory bench tests. Ideally a model is required that is reliant on material properties alone. It is apparent at present that the detail and accuracy of the dynamic modeling far exceeds that of the wear modeling. An alternative approach is that of the Whole Life Rail Model developed by Kapoor et al. (2003). Contacts mechanics predictions are used to determine the plastic flow caused by each wheel passage. This is then used to determine the ratcheting rate. Crack initiation is considered to take place when the strain in a material element exceeds a critical value. Wear takes place when a region (or group of elements) loses integrity. This approach has the advantage that it does not rely on empirical data for life predictions and can implement varying materials properties in the rail section.

22.5 Friction Modification Friction modifiers are applied to the wheel/rail contact to generate required coefficients of friction. These may act to increase or decrease friction depending on the situation. A decrease in friction may be needed where wear rates are high, for example in low radius curves, and an increase may be desired where adhesion loss is prevalent, this would be used in traction and braking. Friction modifiers are divided in to three categories (Kalousec and Magel, 1999): 1. Low coefficient friction modifiers (lubricants) are used to give friction coefficients less than 0.2 at the wheel flange/gauge corner interface. 2. High friction modifiers with intermediate friction coefficients of 0.2 to 0.4 are used in wheel tread/ rail top applications. 3. Very high friction modifiers (friction enhancers) are used to increase adhesion for both traction and braking. Friction modifiers are classified according to their influence after full slip conditions have been reached in the wheel/rail contact, as shown in Figure 22.22 (Eadie et al., 2000). If friction increases after the saturation point the modifiers have positive friction properties, if friction reduces the modifier has negative friction properties. Positive friction modifiers can be described as high positive friction (HPF) or very high positive friction (VHPF), depending on the rate of increase in friction.

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Handbook of Lubrication and Tribology

Positive friction Saturation — full slip

Tractive force (= mN)

Neutral friction

Partial slip Negative friction Free rolling Creep

FIGURE 22.22 Behavior of friction modifiers. (Adapted from Eadie, D.T., Kalousek, J., and Chiddick, K.C., 2000, Proceedings of the 5th International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, pp. 36–41.)

22.5.1 Increasing Friction Friction (or adhesion) loss has a large impact on safety and performance of railway networks. Poor adhesion in braking is a safety issue as it leads to extended stopping distances. If a train experiences poor adhesion in traction when pulling away from a station and a delay is enforced the train operator will incur costs. Similar delays will occur if a train passes over areas of poor adhesion while in service. Work carried out to investigate the causes of adhesion loss using both laboratory and field tests has identified the major causes as being: water (from rainfall or dew), humidity, leaves, wear debris, and oil contamination (Collins, 1972; Broster et al., 1974; Beagley, 1975; Beagley et al., 1975a, 1975b). Work carried out more recently has reemphasized the effect of the problems outlined above and identified further causes of adhesion loss, such as frost and mud deposited on rails by automobile wheels passing over level crossings (Logston, 1980; Nagase, 1989; Jenks, 1997). Most of the work cited above was carried out at relatively low speed. Work on adhesion issues related to high speed lines using both full-scale roller rigs and field measurements has shown that adhesion decreases with train velocity and wheel/rail contact force (Ohyama, 1991; Zhang et al., 2002). The most commonly used friction modifier used on railway networks worldwide to combat adhesion loss is sand. Sanding is used to improve adhesion in both braking and traction. In braking it is used to ensure that the train stops in as short a distance as possible. It usually occurs automatically when the train driver selects emergency braking. Sanding in traction, however, is a manual process. The train driver must determine when to apply the sand and how long the application should last. The sand is supplied from a hopper mounted under the train. Compressed air is used to blow the sand out of a nozzle attached to the bogie and directed at the wheel/rail contact region. This is quite a rudimentary solution and can cause problems to infrastructure. Wheel and rail wear rates increase severely when sand is applied (Kumar et al., 1986; Jenks, 1997; Lewis, 2004b) and there are problems associated with sand build-up at track sites where application is quite frequent. Very high positive friction modifiers to enhance the coefficient of friction from 0.4 to 0.6 are available, but are really only in the development stage. There are a number of different products available, but most involve a solid stick of material that is applied directly to the wheel tread.

22.5.2 Reducing Friction High friction coefficients are most prevalent at the wheel flange/rail gauge corner contact, particularly in curves. Load and slip conditions are also high, which means that high energy consumption and noise generation occurs and wear and rolling contact fatigue are more likely to occur at these sites.

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Industrial Lubrication Practice

22-19

Lubrication is commonly used on curved track to reduce the impact of high loads and slips and reduce friction. This offers a number of benefits including wear reduction (Clayton et al., 1989; Alp et al., 1996; Zhao et al., 1997) and energy saving (Reiff, 1999). The two main methods used to apply lubricant to the wheel/rail contact are wayside lubricators and on-board lubricators. Wayside lubricators are mounted next to the track and apply lubricant to the rail gauge corner. These are of three types: mechanical, hydraulic, and electronic. On-board lubricators apply grease or solid lubricant or spray oil on to the wheel flange, which is then transferred to the gauge corner of the rail. Complex control systems are used in the application process to avoid the application of lubricant at inappropriate locations. Mobile Lubricators, which are essentially railway vehicles designed to apply lubricant to the gauge corner of the track, are also sometimes used. Mechanical wayside lubricators are operated when the wheel makes contact with a mechanical plunger. This operates a pump, which supplies lubricant from a reservoir to a distribution unit. The wheel passes the distribution unit picking up the lubricant and spreading it along the track. Electronic lubricators work slightly differently. They use sensors to detect the approach of a train and activate electric pumps to deliver the lubricant. They are inherently more reliable than mechanical or hydraulic lubricators and can also be adjusted remotely from the track. On-board lubricators supply lubricant to the wheel flange/rail gauge corner. Most apply lubricant to the wheel flange, which is then spread along the rail as the wheel progresses. Some, however, deposit lubricant directly on the rail. Grease or oil spray systems are used, which employ complex control strategies using sensors measuring vehicle speed and track curvature to govern lubricant application. Solid stick lubricators are also available, in which a stick of lubricant is spring loaded against the wheel flange. On-board systems have a number of advantages over wayside lubricators (Sinclair, 2004). There is a reduced safety risk to staff during installation, inspection, and maintenance is easier as it can be carried out in more controlled conditions. The rail will continue to receive some friction control protection in the event of the failure of an individual on-board lubricator. However, despite these advantages, at problem tracks and sites, wayside lubricators are still a necessity. Problems exist with lubrication systems. These are related to both technical and human issues (Thelen and Lovette, 1996). The main technical problems with wayside lubricators are related to mechanical issues such as blocked applicator openings, leaking holes, ineffective pumps, and trigger mechanisms. Poor choice of lubricant can also lead to poor functioning of a lubricator. Human related problems can result from the technical issues. If over lubrication occurs and lubricant migrates onto the rail top, adhesion loss can occur. Train drivers may then be tempted to apply sand to compensate and increase friction, however, this will lead to increased wear and could cause the applicators to become blocked. The thought that application of lubricant will lead to wheel slip can also lead train drivers to switch off on-board lubrication systems. Poor wayside lubrication can lead to potentially serious problems including wheel slip and loss of braking and poor train handling (Roney, 2001). Other issues may be prevention of ultrasonic flaw detection, wastage of lubricant, and high lateral forces in curves and subsequent increase in wear. The key characteristics required of a lubricant are (Roney, 2001): lubricity or the ability of the lubricant to reduce friction; retentivity or the measure of time over which the lubricant retains its lubricity; and pumpability or how easily the lubricant can be applied to the track. The temperature is an issue as some track locations will experience a wide range across which some lubricants may not maintain their pumpability. Some networks use different lubricants in the winter and summer for this reason. The contact temperature is also important as flash temperatures can be as high as 600◦ C to 800◦ C, these can lead to the lubricant in the contact being burned up. Correct positioning of a wayside lubricator is critical to providing effective lubrication. Each site will require something different, which makes this task quite complex. Controlled field testing has been used to assess the reliability and efficiency of wayside lubricators based on a number of factors related to the lubricant including: waste prevention, burn up, distance covered, washing off by rain or snow, and migration to the rail top. This data and factors related to the track, such as length of curve, gradient, and applicator configuration, and traffic, including direction, types of bogie, axle loads, and speeds,

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22-20

Handbook of Lubrication and Tribology Angle a Load Grinding stone

FIGURE 22.23

Facet

Rail grinding.

have been combined to develop criteria and a model for positioning wayside lubricators (de Koker, 1994). Ultimately, however, the most critical element in preserving effective lubrication is maintenance. Once in place wayside lubricators need regular maintenance to prevent the problems outlined occurring.

22.6 Maintenance of Wheels and Rails In order to maintain the required contact conditions between wheel and rail, both are subjected to reprofiling programs. Rail reprofiling is achieved using a grinding process. The objectives of rail grinding are to keep control of the wheel/rail contact conditions (Roney, 2001) to: • maintain a balance between damage mechanisms to negate the need for premature rail replacement • facilitate the desired steering and dynamic stability of railway vehicles • reduce high dynamic loads and track vibrations Rail grinding is performed by specially designed railway vehicles using rotating grinding stones (as shown in Figure 22.23). The flat side of the stones is used, unlike machining process where the edge is used. The material removal process is dependent on the abrasive stone and the load applied to it, as well as the grinding speed and angle of the stone. Rail grinding can serve a number of purposes, as defined by Cooper (1993): 1. Preparative. Cleaning mill scale or surface defects introduced during laying of the rail to ensure a good start to service for the rail. 2. Preventative. Removing layers of fatigued metal before cracking leads to serious damage (see section on Interaction of Wear and Fatigue). 3. Curative. Recovering rail damages during wheel skids. These types of grinding will obviously require different levels of material removal. Wheel reprofiling has the same objectives as rail grinding. It is carried out during routine maintenance. Wheels undergo visual and ultrasonic inspection for damage at set intervals and at the same time regrinding is undertaken depending on the severity of the wear observed or other damage due to RCF or wheel flats etc. Reprofiling is undertaken using special lathes mounted beneath the track. Standards are set for levels of wear. These are related to measurements taken from datums on the flange tip and the inside of the wheel of tread wear and flange thickness reduction, as shown in Figure 22.24.

22.7 Conclusions The wheel/rail interface can appear deceptively simple, but is in fact a complex mixture of many interacting phenomena, which are affected by a multitude of variables many of which are uncontrollable. Management

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Industrial Lubrication Practice

22-21 W

Z Y X

13

70

W X Y Z

Flange thickness (new) Flange thickness (old) Flange height (new) Flange height (old)

Typical values (P8 profile) (mm) 29 24 30 36.5

FIGURE 22.24 Wheel profile limits.

of the wheel/rail interface is therefore a costly business, but is clearly essential for successful and safe rail infrastructure operation. Modeling of the phenomena occurring on wheels and rails leading to degradation and damage is very hard. Most predictive tools are subsequently based on empirical data, which limits their usefulness. The wheel/rail interface is an open system so there are many unpredictable variables and contact conditions vary at virtually every contact point. This means that it would be extremely time consuming to compute wear or fatigue predictions over an entire network. However, there are now models available to predict wear and fatigue of both wheels and rails, which will aid maintenance scheduling. Some railway networks have been slow to take on this new technology, but now at least are starting to appreciate the importance of the role that research in this area can play in improving the management of the wheel/rail interface.

References Alp, A., Erdemir, A., and Kumar, S., 1996, Energy and wear analysis in lubricated sliding contact, Wear, Vol. 191, pp. 261–264. Archard, J.F., 1953, Contact and rubbing of flat surfaces, Journal of Applied Physics, Vol. 24, pp. 981–988. Beagley, T.M. and Pritchard, C., 1975, Wheel/rail adhesion — the overriding influence of water, Wear, Vol. 35, pp. 299–313. Beagley, T.M., McEwen, I.J., and Pritchard, C., 1975a, Wheel/rail adhesion — the influence of railhead debris, Wear, Vol. 33, pp. 141–152. Beagley, T.M., McEwen, I.J., and Pritchard, C., 1975b, Wheel/rail adhesion — boundary lubrication by oily fluids, Wear, Vol. 33, pp. 77–88. Beagley, T.M., 1976, Severe wear of rolling/sliding contacts, Wear, Vol. 36, pp. 317–335. Bogdanski, S., Olzak, M., and Stupnicki, J., 1996, Influence of liquid integration on propagation of rail rolling contact fatigue cracks, Proceedings of the 2nd Mini Conference on Contact Mechanics and Wear of Wheel/Rail Systems, Budapest, pp. 134–143. Bolton, P.J., Clayton, P., and McEwen, I.J., 1982, Wear of rail and tyre steels under rolling/sliding conditions, ASLE Transactions, Vol. 25, pp. 17–24.

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Bolton, P.J. and Clayton, P., 1984, Rolling–sliding wear damage in rail and tyre steels, Wear, Vol. 93, pp. 145–165. Bower, A.F., 1988, The influence of crack face friction and trapped fluid on surface initiated rolling contact fatigue cracks, Transactions of the ASME, Journal of Tribology, Vol. 110, pp. 704–711. British Steel Makers Creep Committee, 1973, BSCC High Temperature Data, The Iron and Steel Institute for the BSCC, London. Broster, M., Pritchard, C., and Smith, D.A., 1974, Wheel–rail adhesion: it’s relation to rail contamination on British railways, Wear, Vol. 29, pp. 309–321. Cannon, D.F. and Pradier, H., 1996, Rail rolling contact fatigue – research by the European Rail Research Institute, Wear, Vol. 191, pp. 1–13. Carter, F.W., 1926, On the action of a locomotive driving wheel, Proceedings of the Royal Society, Vol. A112, pp. 151–157. Clayton, P., Danks, D., and Steele, R.K., 1989, Laboratory assessment of lubricants for wheel/rail applications, Lubrication Engineering, Vol. 45, pp. 501–506. Collins, A.H. and Pritchard, C., 1972, Recent research on adhesion, Railway Engineering Journal, Vol. 1, pp. 19–29. Cooper, J., 1993, Rail flaw detection: a particular challenge, Proceedings of the 5th International Heavy Haul Conference, Beijing, China. Danks, D. and Clayton, P., 1987, Comparison of the wear process for eutectoid rail steels: field and laboratory tests, Wear, Vol. 120, pp. 233–250. Dearden, J., 1960, The wear of steel rails and tyres in railway service, Wear, Vol. 3, pp. 43–49. de Koker, J., 1994, Development of a formula to place rail lubricators, Proceedings of the 5th International Tribology Conference. Eadie, D.T., Kalousek, J., and Chiddick, K.C., 2000, The role of high positive friction (HPF) modifier in the control of short pitch corrugation and related phenomena, Proceedings of the 5th International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, pp. 36–41. Ekberg, A., Kabo, E., and Andersson, H., 2002, An engineering model for prediction of rolling contact fatigue of railway wheels, Fatigue and Fracture of Engineering Materials and Structures, Vol. 25, pp. 899–916. Enblom, R. and Berg, M., 2004, Wheel wear modelling including disc braking and contact environment, accepted for presentation at the 14th International Wheelset Congress, Florida, 17th–21st October 2004. Feller, H.G. and Walf, K., 1991, Surface analysis of corrugated rail treads, Wear, Vol. 144, pp. 153–161. Fries, R.H. and Dávila, C.G., 1987, Wheel wear predictions for tangent track running, Transactions of the ASME, Journal of Dynamics Systems, Measurement, and Control, Vol. 109, pp. 397–404. Garnham, J.E. and Beynon, J.H., 1992, Dry rolling–sliding wear of bainitic and pearlitic steels, Wear, Vol. 57, pp. 81–109. Gao, N. and Dwyer-Joyce, R.S., 2000, The effects of surface defects on the fatigue life of water and oil lubricated contacts, Proceedings of the IMechE Part J, Journal of Engineering Tribology, Vol. 214, pp. 611–626. Grassie, S. and Kalousek, J., 1997, Rolling contact fatigue of rails: characteristics, causes and treatments, Proceedings of the 6th IHHA Conference, Capetown. Grieve, D.G., Dwyer-Joyce, R.S., and Beynon, J.H., 2001, Abrasive wear of railway track by solid contaminants, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transport, Vol. 215, pp. 193–205. Jendel, T., 2000a, Prediction of wheel profile wear — comparisons with field measurements, Proceedings of the International Conference on Contact Mechanics and Wear of Rail/Wheel Systems, Tokyo, Japan, 25–28 July, pp. 117–124. Jendel, T., 2000b, Prediction of Wheel Profile Wear — Methodology and Verification, Licentiate Thesis, TRITA-FKT 2000:9, Royal Institute of Technology, Stockholm, Sweden.

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Industrial Lubrication Practice

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Jenks, C.W., 1997, Improved Methods for Increasing Wheel/Rail Adhesion in the Presence of Natural Contaminants, Transit Co-operative Research Program, Research Results Digest, No. 17. Jiang, Y.Y. and Sehitoglu, H., 1999, A model for rolling contact failure, Wear, Vol. 224, pp. 38–49. Johnson K.L., 1985, Contact Mechanics, Cambridge University Press. Kalker J.J., 1990, Three-Dimensional Elastic Bodies in Rolling Contact, Kluwer, Dordrecht. Kalker, J.J., 1982, Fast algorithm for the simplified theory of rolling contact, Vehicle System Dynamics, Vol. 11, pp. 1–13. Kalousek, J., Fegredo, D.M., and Laufer, E.E., 1985, The wear-resistance and worn metallography of pearlite, bainite and tempered martensite rail steel microstructures of high hardness, Wear, Vol. 105, pp. 199–222. Kalousek, J. and Magel, E., 1997, Optimising the wheel/rail system, Railway and Track Structure, January. Kalousek, J. and Magel, E., 1999, Modifying and managing friction, Railway and Track Structures, May. Kapoor, A., Fletcher, D.I., and Franklin, F.J., 2003, The role of wear in enhancing rail life, Proceedings of the 29th Leeds-Lyon Symposium on Tribology, pp. 331–340. Krause, H. and Poll, G., 1986, Wear of wheel–rail surfaces, Wear, Vol. 113, pp. 103–122. Kumar, S., Krishnamoorthy, P.K., and Prasanna Rao, D.L., 1986, Wheel–rail wear and adhesion with and without sand for a North American locomotive, Journal of Engineering for Industry, Transactions of the ASME, Vol. 108, pp. 141–147. Kumar, S. and Rao, D.L.P., 1984, Wheel–rail contact wear, work, and lateral force for zero angle of attack — a laboratory study, Transactions of the ASME, Journal of Dynamic Systems, Measurement, and Control, Vol. 106, pp. 319–326. Lewis, R. and Olofsson, U., 2004, Mapping rail wear regimes and transitions, Wear, Vol. 257, No. 7–9, pp. 721–729. Lewis, R. and Dwyer-Joyce, R.S., 2004a, Wheel–rail wear and surface damage caused by adhesion sanding, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series No. 43, pp. 731–741. Lewis, R. and Dwyer-Joyce, R.S., 2004b, Wear mechanisms and transitions in railway wheel steels, Proceedings of the IMechE, Part J, Journal of Engineering Tribology, Vol. 218, pp. 467–478. Lewis, R., Dwyer-Joyce, R.S., Bruni, S., Ekberg, A., Cavalletti, M., and Bel Knani, K., 2004, A new CAE procedure for railway wheel tribological design, Proceedings of the 14th International Wheelset Congres, Florida, 17–21 October 2004. Lim, S.C. and Ashby, M.F., 1987, Wear mechanism maps, Acta Metallica, Vol. 35, pp. 1–24. Logston, C.F. and Itami, G.S., 1980, Locomotive friction-creep studies, Transactions of the ASME, Journal of Engineering for Industry, Vol. 102, pp. 275–281. Marshall, M.B., Lewis, R., and Dwyer-Joyce, R.S., 2004, Ultrasonic characterisation of a wheel/rail contact, Proceedings of the 30th Leeds-Lyon Symposium on Tribology, Elsevier Tribology Series No. 43, pp. 151–158. McEwen, I.J. and Harvey, R.F., 1985, Full-scale wheel-on-rail testing: comparisons with service wear and a developing theoretical predictive model, Lubrication Engineering, Vol. 41, pp. 80–88. Nagase, K., 1989, A study of adhesion between the rails and running wheels on main lines: results of investigations by slipping adhesion test bogie, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 203, pp. 33–43. Ohyama, T., 1991, Tribological studies on adhesion phenomena between wheel and rail at high speeds, Wear, Vol. 144, pp. 263–275. Olofsson, O. and Nilsson, R., 2002, Surface cracks and wear of rail: a full scale test and laboratory study, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 216, pp. 249–264. Olofsson, U. and Telliskivi, T., 2003, Wear, friction and plastic deformation of two rail steels — full-scale test and laboratory study, Wear, Vol. 254, pp. 80–93. Pearce, T.G. and Sherratt, N.D., 1991, Prediction of wheel profile wear, Wear, Vol. 144, pp. 343–351. Pyzalla, A., Wang, L., Wild, E., and Wroblewski, T., 2001, Changes in microstructure, texture and residual stresses on the surface of a rail resulting from friction and wear, Wear, Vol. 251, pp. 901–907.

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Reiff, R. and Creggor, D., 1999, Systems approach to best practice for wheel and rail friction control, International Heavy Haul Conference. Roney, M.D., 2001, Maintaining optimal wheel and rail performance, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Sawley, K. and Kristan, J., 2003, Development of bainitic rail steels with potential resistance to rolling contact fatigue, Fatigue and Fracture of Engineering Materials and Structures, Vol. 26, pp. 1019–1029. Sinclair, J., 2004, Friction modifiers, in Vehicle Track Interaction: Identifying and Implementing Solutions, IMechE Seminar, 17th February. Singh, U.P., Roy, B., Jha, S., and Bhattacharyya, S.K., 2001, Microstructure and mechanical properties of as rolled high strength bainitic rail steels, Materials Science and Technology, Vol. 17, pp. 33–38. Stanca, M., Stefanini, A., and Gallo, R., 2001, Development of an integrated design methodology for a new generation of high performance rail wheelsets, Proceedings of the 16th European MDI User Conference, Berchtesgaden, Germany, 14–15 November. Steele, R.K., 1982, Observations of in-service wear of railroad wheels and rails under conditions of widely varying lubrication, ASLE Transactions, Vol. 25, pp. 400–409. Telliskivi, T. and Olofsson, U., 2001, Contact mechanics analysis of measured wheel-rail profiles using the finite element method, Proceedings of the IMechE Part F, Journal of Rail and Rapid Transit, Vol. 215, pp. 65–72. Telliskivi, T. and Olofsson, U., 2004, Wheel–rail wear simulation, Wear, Vol. 257, pp. 1145–1153. Thelen, G. and Lovette, M., 1996, A parametric study of the lubrication transport mechanism at the rail–wheel interface, Wear, Vol. 191, pp. 113–120. Tournay, H., 2001, Supporting technologies vehicle track interaction, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Ueda, M., Uchino, K., Kageyama, H., Motohiro, K., and Kobayashi, A., 1997, Development of bainitic steel rail with excellent surface damage resistance, Proceedings of the 6th IHHA Conference, Capetown. Yokoyama, H., Mitao, S., Yamamoto, S., and Fujikake, M., 2002, Effect of the angle of attack on flaking behaviour in pearlitic and bainitic steel rails, Wear, Vol. 253, pp. 60–67. Zakharov, S., 2001, Wheel/rail performance, in Guidelines to Best Practice for Heavy Haul Railway Operations: Wheel and Rail Interface Issues, International Heavy Haul Association, Virginia, USA. Zhao, X.Z., Zhu, B.L., and Wang, C.Y., 1997, Laboratory assessment of lubricants for wheel/rail lubrication, Journal of Materials Science and Technology, Vol. 13, pp. 57–60. Zhang, W., Chen, J., Wu, X., and Jin, X., 2002, Wheel/rail adhesion and analysis by using full scale roller rig, Wear, Vol. 253, pp. 82–88. Zobory, I., 1997, Prediction of wheel/rail profile wear, Vehicle System Dynamics, Vol. 28, pp. 221–259.

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23 Lubrication in the Timber and Paper Industries 23.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.2 Lubrication in the Timber Industry . . . . . . . . . . . . . . . . . .

23-1 23-2

Timberlands Operations • Felling Equipment • Harvester, Skidder, and Forwarder Lubrication • Chainsaw Lubrication

23.3 Lubrication in the Pulp and Paper Industry . . . . . . . . .

23-5

Wood Yard Operations • Sawmill Equipment Lubrication • Pulp Mill Operations • Pulp Mill Equipment Lubrication • Paper Mill Operations — Wet End Equipment • Paper Mill Operations — Dry End Equipment • Paper Machine Lubrication — Wet End • Paper Machine Lubrication — Dry End • Paper Machine Oil Filterability

23.4 Lubricant Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Paul W. Michael Milwaukee School of Engineering, Fluid Power Institute

23-14

Bearing Lubricant Selection • Gear Lubricant Selection • Synthetic Gear Lubricants • Hydraulic Fluid Selection • Biodegradable Lubricants • Vegetable Oil Based Lubricants — HETG • Synthetic Ester Based Lubricants — HEES • Lubrication Management Strategies • Oil Analysis

References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

23-21

23.1 Introduction Timber harvesting and papermaking push the limits of lubrication technology. In timber harvesting, rugged terrain, extreme temperatures, and the sheer mass of the crop place unparalleled strain upon equipment. In papermaking, the economics of a global market mandate large capital investments, high productivity, and extraordinary equipment reliability. As a result, these industries demand topflight lubricants and lube management practices. Both applications are frequently in close proximity to precious natural resources — wilderness and fresh water. Consequently lubricants for these applications are increasingly formulated to maximize lubricant longevity and minimize environmental impact. This chapter reviews features of lubricant selection, application, and management that are to some extent unique to industries that largely rely on forests for their prime raw material.

23-1

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Handbook of Lubrication and Tribology

23.2 Lubrication in the Timber Industry 23.2.1 Timberlands Operations Timber harvesting systems are characterized by the manner in which wood is moved from the stump to the roadside. The three major types of harvesting methods are tree-length, whole-tree, and shortwood methods [1]. Geographic, economic, regulatory, and silvicultural factors generally dictate the optimum system for timber harvesting. The tree-length harvesting method involves felling, delimbing, and topping trees in the woods and skidding the tree lengths to the landing. The tree lengths are bucked into logs or sticks at the landing. In some instances, the tree lengths are hauled to a mill site for processing. The whole-tree harvesting method involves transporting the entire felled tree to the landing for processing. One advantage of the whole-tree method is that it can utilize the tree’s entire biomass. The shortwood harvesting method involves the conversion of trees into desired length products at the stump, either by hand with chainsaws or by using a mechanized processor which fells, delimbs, and bucks the tree into sawlogs, pulpwood sticks, or other products. The individual pieces are then transported to the landing with a forwarder. The shortwood method is used for thinning a stand or single-tree selection silviculture because it causes less forest damage [2]. Silviculture has features in common with traditional agriculture, particularly with the increasing reliance upon commercial forests and plantation crops such as Eucalyptus. However, the vast majority of wood that is harvested is derived from natural hardwood and softwood forests. Additionally, harvesting timber when the ground is frozen is often desirable because it reduces the impact of heavy equipment on surface vegetation. Thus, timber harvesting equipment must be able to fell, process, and transport a heavy crop through rugged terrain at temperatures that are well below those encountered in most agriculture applications.

23.2.2 Felling Equipment Manual logging is labor intensive and hazardous since it requires workers to operate chainsaws in close proximity to falling trees. In order to reduce labor and insurance costs, the timber industry increasingly relies upon mechanized harvesting equipment such as the feller buncher [3]. A feller buncher is similar to a front-end loader or excavator (see Figure 23.1). It is a rubber tired or tracked vehicle with an articulating extendable arm onto which a felling head is attached. The felling head typically consists of hydraulic grappling devices and a disc saw. Feller bunchers may also be equipped with harvester heads that delimb, debark, and cut trees to length. When felling a tree, the operator positions the felling head at the base of the trunk and hydraulic grappling arms wrap around the tree as the saw removes the tree from the stump (see Figure 23.2). The machine then takes the severed tree (or trees) and lowers it to the ground where it may be loaded into a forwarder or skidded to the roadside. Feller bunchers are advantageous in that they are able to cut trees closer to the ground than chainsaws. However, they may not be suitable for use in steep terrain or for harvesting some large diameter trees. Felled trees generally are transported to a roadside landing area by skidders or forwarders. A skidder is a rubber tired or tracked vehicle that tows cut trees to the landing for transport or processing. Classified as either cable, clambunk, or grapple skidders (see Figure 23.3), their names are derived from the method used to attach the logs to the skidder. They come in a variety of sizes but all have four-wheel drive, good speed, high ground clearance, and diesel power as basic features. Wheeled skidders are used in a wide range of conditions. Track skidders have a bulldozer like undercarriage and are suitable for use in very steep terrain. Skidders are also equipped with an articulating blade which makes them suitable for moving debris that may impede travel through the woods. A forwarder has a front section like a skidder, but instead of a cable winch, has a hydraulic knuckleboom log loader behind the cab and a log deck above the rear axle (see Figure 23.4) [4]. Forwarders are used with conventional chainsaw felling and feller harvesters. Forwarders have a bogie axle that float with the rougher terrain which reduces ground disturbance and counteracts the effect of high center of gravity.

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Lubrication in the Timber and Paper Industries

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FIGURE 23.1 Feller buncher — timber harvesting machine.

FIGURE 23.2 Feller buncher — timber harvesting machine, courtesy TimberPro.

A feller forwarder is a feller buncher with a bunk to the rear of the operator into which the felled trees are lowered and carried to the next tree to be felled. The process is repeated until the bunk is full. The machine then forwards the trees to the landing and unloads them.

23.2.3 Harvester, Skidder, and Forwarder Lubrication Lubrication of timber harvesting machines is similar to that of off-highway construction equipment. Diesel engine oil, hypoid gear lube, transmission oil, hydraulic fluid, and grease are the primary lubricants. The engine oil requirements for this class of equipment are similar to that of construction and off-highway diesel engines. The same is generally true of gear lube and transmission fluid requirements. However, the demands upon lubricants used in timber harvesting equipment are more severe than encountered in many construction applications. Not only are the payloads and temperature ranges extreme, but also construction applications often involve intermittent duty (position and hold) whereas timber harvesting

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Handbook of Lubrication and Tribology

FIGURE 23.3

Grapple skidder — timber harvesting machine.

FIGURE 23.4

Forwarder — timber harvesting machine, courtesy TimberPro.

tends to be continuous duty. These factors create a need for shear stable lubricants with exceptional oxidation resistance and effective antiwear chemistry. Table 23.1 provides a list of lubricant recommendations for timber harvesting equipment.

23.2.4 Chainsaw Lubrication While automated timber harvesting equipment is employed throughout much of Europe and North America, chainsaws remain the most popular tool in the timber harvesting industry. In regions where capital investment is limited, two-cycle chainsaws may be the only power tool available to loggers. Two-cycle engines are used in chainsaw applications because of their light weight, high speed, and ability to be operated at any orientation. They utilize two-cycle engine oils that are diluted at a 50 : 1 fuel to

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Lubrication in the Timber and Paper Industries TABLE 23.1

23-5

Lubrication of Timber Harvesting Equipment

Component Engine Powershift transmission Hydrostatic pump drive Hydrostatic transmission Swing drive Axle differentials (May include wet brakes) Bogie axle Boom hydraulics Harvester heads and grapples Chain auto-lube system

Lubricant

Notes (h = hour)

Diesel engine oil 10W30 or 15W40 Dexron III or C-4 fluid SAE 10 GL-5 Hypoid gear oil 75W90 Multipurpose tractor fluid or multigrade hydraulic oil GL-5 Hypoid gear oil or multigrade hydraulic oil Multipurpose tractor fluid or 75W90 gear oil GL-5 hypoid gear oil 75W90 Multigrade hydraulic fluid HV 32 to HV 68 EP 2 grease 5% Moly Bar and chain oil

Check oil level daily Sample interval: 250 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 1000 h Check oil level weekly Sample interval: 2000 h Check oil level weekly Sample interval: 2000 h Check oil level weekly Sample interval: 2000 h Grease daily Check oil level daily

Source: Schierschmidt, B., personal communication, 2004. Lambert, J., personal communication, 2005.

oil ratio. This 50 : 1 premix is aspirated into the crankcase prior to entering the combustion chamber and serves as both a fuel and a lubricant. Two-cycle engines run very hot, which causes pistons to expand, thus decreasing the piston-cylinder clearance. This loss of clearance increases engine friction and the possibility of scuffing. In order to prevent engine seizure, two-cycle oils incorporate lubricity additives that reduce sliding friction on the cylinder wall. Two-cycle oils must also burn cleanly. Clean burning oils reduce visible smoke exhaust by protecting engines against exhaust port blocking, combustion chamber deposits, ring sticking, and wear. In order to reduce the amount of noncombustible materials entering the combustion chamber, two-cycle oils are formulated with low-ash or ashless additive systems. A typical two-cycle oil contains the following components: • • • •

Synthetic esters for lubricity Solvent for miscibility and low-temperature fluidity Polybutene for lubricity and smoke reduction Ashless antiwear and detergency additives

Bar and chain oils are also used in chainsaw applications. Typically these products are SAE 10 or SAE 30 lubricants that contain tackifiers to enhance adhesion. They reduce groove and rail wear on bars by minimizing drive link friction. Since bar and chain oils are “total loss” lubricants, that is they are used once and then released into the environment, increasingly biodegradable lubricants are used in this application [5].

23.3 Lubrication in the Pulp and Paper Industry The pulp and paper industry presents a distinctive environment for lubricants and maintenance engineers. Intense global competition and environmental regulation create an economic environment that favors large-scale, highly capitalized operations. In order to generate a satisfactory return on investment (ROI) in this economic climate, it is necessary to maintain high levels of equipment reliability and productivity.

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23-6

Handbook of Lubrication and Tribology Veneer logs Plywood mill

Wood residues

Pulpwood

Particle board mill

Sawlogs Sawmill

Pulpwood

Chemical pulping

Pulpwood

FIGURE 23.5

Wood residues

Paper mill

Groundwood mill

Integrated wood-yard operations.

23.3.1 Wood Yard Operations Wood yard operations in lumber and pulp mills have a number of similarities. In both industries, logs are unloaded from semi-trailers or railcars and transported to log decks by modified end loaders or knuckle-boom cranes. If necessary, tree length logs are cut to length on slasher decks. Saw logs are then conveyed to debarking in order to produce clean lumber or pulp. Rosser-head or ring debarkers are used in sawmill applications while drum debarkers are used in pulp mills. Drum debarkers are not used in sawmill applications because they cause the log ends to splinter, which reduces their value as lumber. In pulp mill operations debarked wood is chipped, washed, and screened. In sawmill operations debarked wood is sawed, edged, trimmed, and kiln dried. Sawmill trimmings are processed through a chipper for use in pulp mills. In both applications, bark is usually ground into small pieces by refuse hogs and used as boiler fuel. In order to maximize economic utilization of timber resources, lumber and pulping operations may be integrated as depicted in Figure 23.5.

23.3.2 Sawmill Equipment Lubrication Sawmills utilize a vast assortment of kickers, turners, roller decks, and chain conveyors. Chain oil, gear lube, and grease are the primary lubricants used in this equipment. In order to avoid staining the wood, light colored oils are preferred for the lubrication of roller-chain conveyors. Optimizing solid-wood yields is particularly important in sawmill operations where laser sensors and hydraulic setworks are used to maximize the quantity and value of lumber extracted from logs. Band saws are increasingly used in these applications because they have a narrow kerf that produces about half the sawdust generated by a circular saw. While sawdust is a useful fuel source for the boilers that heat lumber kilns, solid wood has a much higher economic value. With the periodic exception of kiln fan bearings, sawmill operating temperatures are not particularly high. However, low temperatures, moisture, dirt, and sawdust are endemic, as is shock-loading due

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Lubrication in the Timber and Paper Industries

23-7

to the constant pounding of the logs that are processed. These conditions mandate frequent grease application, use of synthetic gear oils, and generous lubrication of roller-chain conveyors. Table 23.2 provides lubrication recommendations and guidelines for sawmill equipment.

23.3.3 Pulp Mill Operations Pulp is the key raw material for papermaking. Pulp is made by mechanically or chemically separating the fibers in wood or other cellulosic materials from nonfibrous material. Mechanical pulping converts wood into pulp by shearing the fibers in logs or chips with grinders and disk refiners. Chemical pulping converts wood chips into pulp by cooking them in an aqueous solution at high temperatures and pressures. The two principal chemical pulping methods are the (alkaline) kraft process and the (acidic) sulfite process. In the kraft process wood chips are cooked in a solution of sodium hydroxide (NaOH) and sodium sulfide (Na2 S). In the sulfite process wood chips are cooked in a solution of sulfurous acid (H2 SO3 ) and bisulfite ion (HSO− 3 ). Both processes chemically degrade the lignin that binds the fibers within the wood. Due to advantages in chemical recovery and fiber quality, the kraft process has gained larger market acceptance. In an integrated mill, pulp is usually stored as a medium consistency stock prior to forming, pressing, and steam-cylinder drying on the paper machine. For nonintegrated pulp and fiber operations, pulp must be dewatered to decrease transportation costs. Dewatering is usually performed on a lapping press, which utilizes a sheet forming wet end, a press section, and a drying section similar to a paper machine [6]. The laps that are produced by this machine are stacked and bailed by a hydraulic layboy for economical transportation by truck or rail.

23.3.4 Pulp Mill Equipment Lubrication Lubrication of pulp mill equipment combines demanding elements of off-road construction and industrial equipment lubrication. Heavy loads, wide ranging temperatures, and contamination challenge the effectiveness of lubricants. High capital equipment costs and global competition also make equipment reliability and productivity a priority. Thus effective lubricants and lubrication management systems are critical to the economic success of mill yard operations. Use of synthetic lubricants and consolidation of lubricants are often helpful means of improving the efficiency of mill yard equipment lubrication. Table 23.3 provides a list of lube recommendations for pulp mill equipment.

23.3.5 Paper Mill Operations — Wet End Equipment Papermaking furnish is prepared in the Approach System of the paper machine. The machine chest, fan pump, headbox, and slice are the key components of the approach system. Here, pulp is diluted, mixed, and screened prior to being pumped into the headbox by the fan pump. Pressurized stock from the fan pump is discharged from the headbox through the slice opening and deposited on the forming fabric or wire. The wire is a continuous plastic mesh belt that carries the high-water content stock across a series of forming boards on the Fourdrinier. Stock is dewatered on the Fourdrinier by hydraulic pressure gradients created by these forming boards or foils. There also is a series of vacuum boxes on the Fourdrinier that assist in water removal. The forming section of the paper machine, which includes the Fourdrinier (Figure 23.6), incorporates several types of rolls: • The breast roll, located below the slice opening, where stock is deposited onto the forming fabric or wire • The dandy roll, positioned near the end of the Fourdrinier, which restructures fibers within the web and reduces the water content at the top of the sheet • The suction couch roll, at the end of the Fourdrinier, which transfers the formed sheet into the press section of the paper machine • Wire turning, guide, and tensioning rolls that control forming fabric travel

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23-8 TABLE 23.2

Handbook of Lubrication and Tribology Lubrication of Sawmill Equipment

Machine Log decks — kickers, loaders, and turners

Rosser head debarker

Component

Circle saw

Lithium complex grease NLGI 2 HM 32 HM 32

Chain drive gearbox

Synthetic ISO 220

Bull wheel bearings

Grease daily

Chains Chain drive gearbox

Lithium complex grease NLGI 2 Lithium complex grease NLGI 2 HM 32 Synthetic ISO 220

Hydraulic system

HM 32

Chains and air cylinders Chain drive gearbox

HM 32 Synthetic ISO 220

Arbor bearings

Lithium complex grease NLGI 2 HM 32

Check weekly Sample quarterly Hand oil weekly Inspect weekly Sample annually Grease weekly

Hydraulic turn, set, and feed

Band saw

Grease weekly Hand oil weekly Check weekly Sample quarterly Inspect weekly Sample annually

Grease daily Hand oil weekly Inspect weekly Sample annually

Check weekly Sample quarterly Hand oil weekly Grease weekly

Carriage knees, slides, and pinion Cable drum and idler bearings

HM 32 Lithium complex grease NLGI 2

Band saw wheel bearings

Hydraulic turn, set, and feed

Lithium complex grease NLGI 2 Lithium complex grease NLGI 2 HM 32

Saw guides and slides Wheel yoke screw drives

Spindle oil ISO VG 10 Synthetic ISO 220

Airline oiler Yoke bushing oil cup

HM 32 HM 32

Arbor bearings

Lithium complex grease NLGI 2 Spindle oil ISO VG 10

Grease weekly

Press and feed roll bearings

Gang saw

Notes

Conveyor and hour glass roll bearings Chains and air cylinders Log turner hydraulic system

Head wheel bearings

Ring debarker

Lubricant

Blade guide auto lube system

Clean and repack annually Grease weekly Check weekly Sample quarterly Check weekly Inspect weekly Sample annually Check weekly Fill weekly

Check weekly

Chipper

Main bearings

Lithium complex grease NLGI 2

Clean and repack at 6 month interval

Chip shaker screen

Eccentric bearings

Lithium complex grease NLGI 2 Lithium complex grease NLGI 2

Grease weekly

Knuckle bushings

Grease weekly

Bark hog

Main bearings

Lithium complex grease NLGI 2

Grease monthly

Kiln

Blower bearing fittings

Lithium complex grease NLGI 2 Synthetic ISO 220

Grease monthly

Blower bearing oil cups Source: Cornell, S., personal communication, 2005. Leja, R., personal communication, 2005.

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Fill weekly

Lubrication in the Timber and Paper Industries TABLE 23.3

23-9

Lubrication of Pulp Mill Equipment

Machine Debarker feed and deicing deck Drum debarker

Component

Lubricant

Notes

Hydraulic system

HV 46

Motor Gear coupling Fluid coupling

Grease Grease HM 68

Chain gear drive Chain Trunnions — grease lubed Trunnions — oil lubed

Synthetic ISO 220 HM 68 Lithium complex #2 Syn 460

Hydraulic discharge gate

HM 68

Chipper

Synchronous motor Bearings

Grease Grease

Motor bearing grease Lithium complex NLGI #2

Belt and screw chip conveyors

Motor PIV coupling Gear coupling Reducer

Grease Grease Grease Synthetic ISO 220

Chains Screw bearings Rollers Tail head

HM 68 Grease Grease Grease

Apply motor bearing grease annually Monthly Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly Monthly Monthly Monthly

Falk drive

Synthetic ISO 220

Chains Screen bearings

HM 68 Grease

Pump bearings

HM 68

Gear coupling Reducer

Grease Synthetic ISO 220

Agitator bearings

Grease

Main drive shaft Plate adjust gear case

Grease Synthetic ISO 460

Rotary and disc chip screens

Digestors and agitators

Refiner

Washer

Lime kiln

Lube pump bearings

Grease

Motor bearings Coupling Gear case

Grease Grease Synthetic ISO 460

Shaft

Grease

Motor Coupling Reducer

Grease Grease Synthetic ISO 460

Drive gears Trunnions — grease lubed Trunnions — oil lubed

Open gear compound Grease Synthetic ISO 460

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Inspect weekly 3-month sample interval Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly 3-month sample interval Automatic drip or brush oiler Grease monthly Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval

Inspect weekly 6-month sample interval Check oil level weekly Lithium complex NLGI #2 Inspect daily 6-month drain interval Apply coupling grease annually Inspect weekly 6-month sample interval Monthly Monthly Inspect weekly 6-month sample interval Monthly Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Monthly Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 3-month sample interval Inspect weekly Grease weekly Inspect weekly 6-month sample interval

23-10 TABLE 23.3

Handbook of Lubrication and Tribology Continued

Machine Boiler

Component

Lubricant

Notes

Motor bearings

HL 32

Fluid coupling

HL 32

Feed water pump

HL 32

Steam turbine

Reservoir

HL 32

Inspect weekly 3-month sample interval

Stock and chemical pumps

Motor bearings Coupling Reducer Pump bearings

Grease Grease Synthetic ISO 460 HM 68

Apply motor bearing grease annually Apply coupling grease annually Inspect weekly 6-month sample interval Inspect weekly 6-month drain interval

Whitewater pumps

HM 68

Vacuum pumps

HM 68

Double wire press drive reducer Dancer roll Hydraulic felt tensioner

Synthetic ISO 220 Grease HM 46

Inspect weekly 3-month sample interval Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Grease quarterly Inspect weekly 6-month sample interval

Convey, lift, and discharge reducers Up-ender and scissors lift hydraulics Shafts and rolls

Synthetic ISO 220 HM 46

Lapping press

Pulp layboy

Grease

Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval

Inspect weekly 6-month sample interval Inspect weekly 6-month sample interval Grease quarterly

Source: Sobczak, T. and Prescher, S., personal communication, 2004.

Paper is transferred from the suction couch roll to the suction pickup roll as it enters the Press Section of the machine. In the press section of the machine, the paper web is conveyed through a series of two-roll nips by a synthetic felt. Pressing the paper removes water from the sheet and consolidates the web. Since it is more economical to remove water from paper by pressing than evaporation, press section design and operation is carefully optimized for maximum water removal. A myriad of specialized rolls are used to dewater paper in the press section including: • • • •

Perforated suction press rolls with suction boxes that draw water through the felt Grooved and blind-drilled rolls that provide a path for water to escape the nip Swimming rolls that utilize oil pressure to compensate for roll deflection Nipco rolls with individual hydrostatic pistons that allow precise crown control in different zones of the roll • Extended nip and other stationary shoe presses that are used to increase nip dwell times • Granite and covered steel rolls that serve as sealing surfaces when mated with controlled crown and perforated rolls • Felt turning, guide, and tensioning rolls that control press fabric travel

23.3.6 Paper Mill Operations — Dry End Equipment After press section operations, paper is transferred to the Dryer Section of the machine where it is carried by a dryer felt over a series of rotating steam-heated cylinders or dryer rolls. Typically, the dryer section of

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Lubrication in the Timber and Paper Industries

23-11 Dryer section

Headbox

FIGURE 23.6

Fourdrinier

Calender

Press

Reel

Typical paper machine layout, adapted from Smook courtesy Angus Wilde Publications [6].

the machine contains 40 to 100 dryer rolls that are 1 to 2 m in diameter. Water is removed from the paper in a two-step process as it is heated by contact with the rotating cylinder and flashes off as steam in the open draw or pocket between upper and bottom cylinders [7]. Pocket ventilators are utilized to hasten evaporation by supplying hot dry air through the felt rolls or an exterior duct. In the manufacture of lightweight paper such as tissue a single large dryer cylinder, known as a Yankee dryer, is employed. Yankee dryers vary in size from 4 to 7 m in diameter. A pressure roll transfers the sheet to the Yankee dryer surface. As the thin sheet rotates around the Yankee, it quickly dries and then is scraped off by a metal blade known as a doctor. Following the last dryer, paper calendaring and reeling operations are performed. The calendar is a set of rolls that smooth the paper surface to improve printability. In traditional hard-nip machines the paper is compressed to a uniform thickness by two hard rolls. In super calendars, mating a hard roll with a soft roll forms the nip. Calendaring can also be performed off-machine, although it is less common. The finishing section contains the reel and winder. The reel collects the paper coming off the paper machine. The winder is used to cut the paper into smaller rolls that are then shipped to their customers or gets converted onsite into the end product.

23.3.7 Paper Machine Lubrication — Wet End The wet end of a paper machine operates in a deluge of water and fiber. It is also subject to frequent wash-downs with high-pressure aqueous cleaners. These conditions present a significant challenge in terms of bearing lubrication. Upper press rolls are particularly problematic because of the position of the load zone [8]. Since grease can provide a physical barrier that inhibits water ingression, it is often the preferred lubricant for wet-end bearings [9]. Lubricating grease is a solid to semi-fluid substance formed by mixing a gelling agent or soap thickener into a lubricating fluid. The thickening agent keeps the lubricant in place, while a combination of heat, mechanical shear, and capillary action transfers lubricating fluid to bearing surfaces [10]. Typically NLGI #2 calcium sulfonate or lithium complex grease with good water washout and corrosion resistance is used for wet-end bearing lubrication. Since water contamination may cause grease to de-gel and soften in consistency, evaluating the roll-stability of grease after addition of water is advisable [11]. As shown in Table 23.4, wet-end bearings may also utilize oil. While grease offers the benefit of sealing out contaminants, oil is more effective at contamination removal. In high speed machines, lubricating oil systems with coolers, filters, and vacuum dehydrators are often employed. This is necessary because contamination removal is critical in bearings that rotate at high peripheral velocity [12].

23.3.8 Paper Machine Lubrication — Dry End Whereas some older machines with grease lubricated journal bearings still produce paper and liner board, oil lubricated spherical roller bearings are used in the dryer sections of modern machines. Since high-pressure steam passes through the journal of a dryer drum, bearing temperatures are quite high,

© 2006 by Taylor & Francis Group, LLC

23-12

Handbook of Lubrication and Tribology TABLE 23.4

Lubrication of Paper Machine Equipment

Machine

Component

Lubricant

Gould pump

Sight glass oiler

HM 68

Gorman Rupp pump

Sight glass oiler

HM 68

Sun Flo pump

Reservoir

Fan pump

Sight glass oiler

Synthetic 5W-30 HM 68

Pulper agitators

Reducer

PMO 220

Headbox slice

Reducer

Fourdrinier wire turning roll Couch roll

Drive gearbox

Synthetic ISO 220 Synthetic 1ISO 220

Drive gearbox

Synthetic ISO 220

Bearings grease lubricated Bearings oil lubricated

Grease

Wet end Wet end

PMO 220

Wet end press section Profile roll

Drives

Synthetic ISO 220

Hydraulic system

HM 100

Dryer bearings

Lube system

PMO 220 or PMO 320

Dryer planetary or hypoid drives

Lube system

PMO 150, 220, or 320

Size press Reel

Worm reducer Hydraulic system

Synthetic ISO 460 HM 46

Lineshaft turbine

Circulation system

HL68

Doctor oscillators Nash vacuum Pump

Reducer Reservoir

Synthetic ISO 460 HM 100

Hood lifts Disc filter Exhaust fans

Reducer Worm reducer Reducer Fan bearings

Exhaust fans

Motor bearings

Synthetic ISO 460 Synthetic ISO 460 Synthetic ISO 220 GC/LB Grease Grease

Transfer pumps

Falk reducer

PMO 220

Winder/unwinder

Hydraulic system

HM 46

Controlled crown roll

Hydraulic system

HM 46

© 2006 by Taylor & Francis Group, LLC

Notes Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect daily 6-month drain interval Inspect weekly 6-month sample interval 6-month oil sample interval Inspect daily 12-month sample interval Inspect daily 12-month sample interval Grease monthly with a tacky lithium complex NLGI 2 Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval Inspect daily 12-month sample interval Inspect daily 6-month oil sample interval Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval Inspect sight glasses for oil flow and clarity daily Monthly oil sample interval 12-month drain interval Inspect daily 12-month sample interval Inspect weekly 6-month sample interval 12-month drain interval Inspect weekly 6-month oil sample interval 12-month drain interval 12-month drain interval 6-month oil sample interval Grease quarterly Apply motor bearing grease annually Inspect weekly 6-month oil sample interval Inspect daily 6-month oil sample interval Inspect daily 6-month oil sample interval

Lubrication in the Timber and Paper Industries TABLE 23.4

23-13

Continued

Machine

Component

Controlled crown roll

Circulating system

Reel

Drum drive

Calendar

Nip loading hydraulics

Lubricant

Notes

Synthetic ISO 220 Synthetic ISO 220 HM 68

6-month oil sample interval 6-month oil sample interval Inspect daily 6-month oil sample interval

Source: Ahrens, J.F. and Schlaefer, P., personal communication, 2005.

TABLE 23.5 Pall Filterability Index Test Criteria for Paper Machine Oils Volume passed, ml

Filter length of service

4000 2000 1000

Excellent Acceptable Short

especially when the journals are not insulated [13]. Hooding of the dryer section, which increases bearing exposure to radiant heat from the dryer cans, further exacerbates conditions for dryer bearings [14]. Consequently paper machine oils must be formulated to resist high temperature degradation and associated bearing deposits. Dryer bearing lubricants must be formulated to tolerate water contamination because steam and condensate leaks are a familiar occurrence. In order to prevent bearing corrosion and fatigue as well as avoid an unscheduled machine shutdown for high-water levels, it is common practice to routinely drain water from the bottom of paper machine oil reservoirs. Use of vacuum dehydrators is also a common strategy for coping with water contamination [15]. Since rapid water separation reduces the demands upon vacuum dehydrators and decreases the likelihood of filter by-pass due to additive precipitation, demulsibility is an important characteristic of paper machine oil.

23.3.9 Paper Machine Oil Filterability The filterability characteristic of a fluid can be defined as its ability to pass through a filter without giving rise to undue pressure drop. Excessive pressure drop is undesirable because it can lead to abbreviated filter life. Normally, flow degradation occurs over a period of time as filters accumulate dirt, sludge, and wear debris. When filters become blocked by additives that precipitate out of oil as a result of a chemical reaction with water or other liquid contaminants, filter usage and replacement costs can skyrocket. Several test methods have been developed for evaluating the filterability of lube oils. Filterability tests generally consist of filtering a specified quantity of fluid through a standard medium. The results are typically reported in terms of a ratio between flow rates with and without water in an attempt to compensate for the effect of viscosity on filterability [16]. However, filterability tests designed with hydraulic fluids in mind are generally not well suited for high-viscosity paper machine oils. For instance, ISO 13357-1 : 2002, Petroleum Products — Determination of the filterability of lubricating oils, is limited to a viscosity of up to ISO VG 100 [17]. The Pall Filterability Index for Paper Machine Oils test was developed to evaluate high-viscosity oils [18]. In this test 4000 ml of oil is heated to 145◦ F for 1 h and then pumped through a 3-µm nylon filter membrane. The amount of fluid that can be filtered before reaching a terminal pressure drop of 25 psi is measured. The test is performed with and without the addition of water. By measuring the total volume of fluid filtered rather than the time necessary to filter a specific volume of oil, the effect of fluid viscosity is reduced. The Pall Filterability Index rating’s relationship to filter life is shown in Table 23.5.

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology

23.4 Lubricant Selection 23.4.1 Bearing Lubricant Selection Roller bearings and journal bearings are lubricated by different mechanisms. In a journal bearing, a wedge of oil is drawn into the load zone as the shaft rotates. Hydrodynamic pressure created by the lubricant as it is drawn between these converging surfaces produces an oil film barrier that supports the bearing load. Since the journal and bearing surfaces conform to each other, the bearing load is distributed through the lubricant film over a relatively large area. In an antifriction or roller bearing, oil adsorbed on converging surfaces is also drawn into the bearing load zone by hydrodynamic forces. However, the contact surfaces of a bearing roller and race are nonconformal and the load is concentrated at the point or line where they make contact. The high pressure created within these contacts cause oil to undergo an instantaneous exponential viscosity increase. At the same time, surfaces elastically deform under load creating a Hertzian contact zone. As a result, the viscous lubricant is trapped within the bearing load zone and prevents surface contact. This enables a very thin elastohydrodynamic film to support the high loads generated within an antifriction bearing. Pressure distribution within an EHD contact is illustrated in Figure 23.7 [19]. The minimum film thickness, hm , occurs at the outlet of the conjunction zone. The value of hm , in a roller bearing can be determined using Equation 23.1 [20]. Bearing film thicknesses are normally in the 0.5 µm range [21]. This equation shows that film thickness is most sensitive to surface velocity (u), viscosity (µ0 ), and the viscosity–pressure coefficient (α) of the lubricant. The influence of the elastic modulus (E) and load (w) is very small since an increase in load merely increases the size of the Hertzian contact zone. As with conventional hydrodynamic lubrication, the lubricant’s viscosity must be high enough to resist being squeezed out of the converging zone. At the same time the entrainment velocity must be sufficient to continuously replenish the oil film. Generally the EHD film thickness should be 1.5 to 4 times the composite surface roughness to prevent contact of surface asperities. Failure to generate an adequate EHD film will result in a reduced bearing fatigue life [22].

hm = 2.65

(µ0 u)0.7 α 0.54 R 0.43 E 0.03 w 0.13

(23.1)

hm = film thickness at the rear constriction µ0 = viscosity at atmospheric pressure α = pressure–viscosity coefficient u = velocity defined as u = 12 (u1 + u2 ) where u1 and u2 are the individual velocities of the moving surfaces R = the radius of equivalent cylinder w = load per unit width E = elastic modulus if equivalent cylinder (flat surface assumed completely rigid) A variety of simplified methods have been developed for determining the minimum viscosity requirements for antifriction bearings [23,24]. An approximation of the oil viscosity required to achieve acceptable oil film thickness at the bearing operating temperature can be obtained by solving Equations 23.2 or 23.3, where n is the rotational speed in rpm and dm is the mean bearing diameter in mm [25]. The mean bearing diameter equals the average diameter of the bearing OD and ID (dm = (D + d)/2). where

4500
1000 1/3 v1 = √ n ndm 4500 v1 = √ ndm

© 2006 by Taylor & Francis Group, LLC

for n < 1000 r/min

for n ≥ 1000 r/min

(23.2) (23.3)

Lubrication in the Timber and Paper Industries

23-15

Hertzian pressure

Pressure Inlet region (Pumps film up)

hm Hertzian region (Rides it)

Outlet region (discharges it)

FIGURE 23.7 Pressure distribution within an EHD contact [18].

When applying these equations to paper machine dryer bearings, commonly the viscosity specified by the machine manufacturer is lower than the optimum level for full-film EHD lubrication. This is because process and design modifications that increase machine productivity also tend to cause higher bearing temperatures. When these conditions occur it is desirable to increase oil viscosity to provide enhanced bearing life. Whether or not this is feasible often depends upon the capacity of return lines to drain the oil back to the machine reservoir.

23.4.2 Gear Lubricant Selection The pulp and paper industry’s use of industrial gear drives is extensive. A recent survey of a single North American pulp mill identified 816 individual industrial gear drives [26]. In a nearby paper mill, 335 gear drives were counted among the support equipment for a single paper machine [27]. The majority of the gears used in these applications are enclosed gear drives of the spur, helical, and worm configuration. The contact mechanics of a spur, helical, or bevel gear creates a combination of sliding and rolling friction. When a gear tooth enters the mesh, contacting teeth slide toward each other with a slight rolling action. Sliding diminishes to pure rolling as pitch lines intersect, after which the contacting gear surfaces slide away from each other. Because of the complexity of this interaction, gears can operate under three different modes of lubrication: boundary, mixed, and elastohydrodynamic. Equation 23.1 may be used as a generalized form for calculation of the EHD film thickness in a line contact. In order to extend this equation to gear applications it is necessary to incorporate the mesh angle and transmission ratio, which adds to the complexity of the calculation [28]. An estimate of the required viscosity can be made based upon pitch line velocity using Equation 23.4 [29]. υ40 =

7000 (V )0.5

(23.4)

υ40 = lubricant kinematic viscosity at 40◦ C, cSt. V = operating pitch line velocity, ft/min. V = 0.262dn. d = operating pitch diameter of pinion, in., n = pinion speed, rpm. At low pitch line velocities, it is difficult to generate an effective EHD film. When low speeds are combined with high loads, boundary lubrication prevails. In boundary lubrication, the average film thickness is less than the surface roughness and asperity contact occurs [30]. As a result, friction and wear are dominated by contact surface properties, rather than bulk fluid properties. Extreme pressure (EP) additives are frequently incorporated in gear lubricants for the purpose of reducing friction and wear where

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Handbook of Lubrication and Tribology

TABLE 23.6 Properties of Synthetic and EP Gear Lubricants Property

Test procedure

EP gear oil

Synthetic gear oil

Viscosity index

ISO 2909 ASTM D2270

90 minimum

120 minimum

Oxidation stability

ASTM D2893

10

25 µ wet or dry

25 µ wet or dry

Demulsibility

Load carrying property

Filterability

ASTM D2711 MOD Max. percent water in the oil after 5-h test Max. cuff after centrifuging Min. total free water collected ASTM D2782 Timken test DIN 51 354 FZG test None

under boundary conditions. Typically sulfur and phosphorus compounds serve as EP additives in mineral oil based gear oil formulations. These additives modify the gear tooth surface by forming a tenacious reaction film that prevents scuffing, metal adhesion, and wear [31]. In 1994 the AGMA 250.04 Standard for Industrial Gear Lubrication was replaced by the ANSI/AGMA 9005-D94 American National Standard for Industrial Gear Lubrication [32]. This standard provides lubricant classifications and maintenance guidelines for industrial gearing. At the time of standard development, specifications for EP oils were upgraded to reflect advancements in additive technology. In addition, specifications for synthetic lubricants were incorporated in the new standard. These standards are comparable. However, the load carrying requirements are more demanding for EP mineral oil lubricants than synthetics (Table 23.6).

23.4.3 Synthetic Gear Lubricants Use of synthetic gear lubricants is popular in pulp and paper applications because they can reduce maintenance costs and improve equipment reliability. Synthetic gear lubricants are more oxidation resistant and thermally stable than mineral oil gear lubricants. They also maintain a stable viscosity over wider temperature range. Consequently synthetic lubricants provide the opportunity for extended drain intervals and generate more effective hydrodynamic and elastohydrodynamic lubricating films. Improved energy efficiency may also be realized through the use of synthetic lubricants, especially in worm drive applications (Table 23.7). Relative to mineral-derived lubricants, synthetics have a greater film thickness at high temperatures due to their higher viscosity index [33]. In rolling element bearings this can result in a fourfold increase in bearing life [34]. Polyalphaolefin (PAO) and polyalkyleneglycol (PAG) are the most popular synthetic base stocks for formulating gear lubricants. PAG is synthesized by polymerizing ethylene oxide with propylene oxide. PAO is synthesized by polymerizing C-10 or C-12 alphaolefins. Because these molecules are synthetically derived, they are devoid of wax molecules that hinder low-temperature fluidity and sulfur compounds that compromise oxidation stability [35].

© 2006 by Taylor & Francis Group, LLC

Lubrication in the Timber and Paper Industries

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TABLE 23.7 Comparison of Efficiency of Lubricants in Worm Gear Applications Lubricant ISO 460 Mineral oil ISO 460 Polyalphaolefin ISO 460 Polyalkylene glycol

Viscosity index

Percent efficiency (%)

95 >150 >220

60 68 78

Source: Lauer, D.A., Gear Solutions, August 2004, pp. 21–30.

TABLE 23.8 Typical Test and Service Life of Solvent Refined and Hydrocracked Base Stocks Typical test and service life (h = hours, yr = years, m = months)

Turbine oil, TOST life Steam turbine, service life Hydraulic oil, TOST life EP gear oil, service life

Solvent-refined group I

Hydrocracked group II and III

4000 h 10 yr 2000 h 6m

18000 h 25 yr 6000 h 1 yr

Some of the benefits of synthetic lubricants can be obtained by using hydrocracked base stocks. As can be seen from Table 23.8, the oxidation life of lubricants formulated with these base stocks is at least double that of solvent refined lubricating oils [36]. Extreme Pressure (EP) gear oils exhibit less of a benefit from the use of hydrocracked base stocks because sulfurized EP additives catalyze oil oxidation. For this reason most synthetic gear oils formulated for industrial applications do not incorporate sulfurized EP additives.

23.4.4 Hydraulic Fluid Selection Viscosity is one of the most important criteria in the selection of a hydraulic fluid. A hydraulic fluid that is too low in viscosity will cause low volumetric efficiency, fluid overheating, and increased pump wear. A fluid that is too high in viscosity will cause poor mechanical efficiency, difficulty in starting and wear due to insufficient oil flow. Selecting the proper viscosity fluid requires an understanding of the viscosity requirements of the hydraulic components as well as the operating temperature range. Hydraulic component manufacturers were surveyed regarding the fluid viscosity requirements of their pumps and motors [37]. The majority of equipment was found to provide satisfactory performance with an operating viscosity range of 13 to 860 cSt. Based upon this viscosity range, a Temperature Operating Window (TOW) chart was developed for straight grade hydraulic fluids shown in Figure 23.8. When selecting a hydraulic fluid using TOW criteria, determine the lowest ambient temperature at start up and the highest temperature in use. Any fluid that has a TOW that encompasses this range may be used in the application. Since most of the hydraulic systems within a paper mill incorporate coolers and thermostats, operating temperatures are relatively stable. Consequently it is often possible to consolidate within a paper mill to a single VG 46 or ISO VG 68 hydraulic fluid. In forestry and wood yard applications, the operating temperatures of hydraulic systems are wideranging. Often the use of straight-grade oil is only feasible if seasonal oil changes are performed [38]. Unfortunately it is difficult to thoroughly drain a hydraulic system on a timber harvester or a log loader because upwards of half the fluid can remain trapped within lines and cylinders. As a result the seasonal oil change strategy often yields a mixture of viscosities that is inadequate at low and high temperatures. Multigrade hydraulic fluids may be used in forestry and wood yard operations to eliminate the need for seasonal oil changes. Multigrade oils contain a viscosity index improver that enhances a fluid’s resistance to viscosity change due to temperature. As a result, multigrade fluids have good low-temperature pumpability

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Handbook of Lubrication and Tribology

110

230

100

212 94

90 80

Temperature°C

60

140

64 55

50 40 30

122 104

44

86

32

20

68 +10

10

+4 –2

0

–30 –40

14

–15 –23

–20

–4

–33 10

50 32

–8

–10

FIGURE 23.8

158

73

Temperature°F

70

194 176

84

–22 15

22

32

46

68

100

–40

TOW for 13 to 860 mm2 /sec (cSt), 100 VI hydraulic fluid.

and good high temperature lubricity. Superior viscosity stability of multigrade oils also reduces mechanical energy losses during start-up and volumetric efficiency losses at high temperatures, which can result in enhanced productivity and fuel savings [39]. Selecting the optimum multigrade oil for these applications may be done using an ASTM Viscosity– Temperature Chart using the following procedure: 1. 2. 3. 4. 5. 6.

Determine the minimum and maximum viscosity requirements for system pumps and motors. Estimate the lowest and highest anticipated fluid temperatures in operation. Plot the highest recommended viscosity at the lowest anticipated temperature. Plot the lowest recommended viscosity at the highest anticipated temperature. Draw a line connecting these points. The viscosity requirements of the fluid are defined by the intersection of the line at 40 and 100◦ C.

23.4.5 Biodegradable Lubricants Biodegradable lubricants are increasingly used in forestry applications because of concern over the possibility of lubricants contaminating the environment, especially waterways. Most biodegradable lubricants exhibit two key environmental characteristics: virtual non-toxicity to aquatic life and aerobic biodegradability. In addition, oils that are derived from soybean, rapeseed, sunflower, and other plants are a renewable resource. Since these plants utilize carbon dioxide during photosynthesis as shown in Figure 23.9, they also do not increase green-house gasses in the atmosphere [40]. Organizations such as the Organization for Economic Co-operation and Development (OECD), the Co-ordinating European Council (CEC), and the U.S. Environmental Protection Agency (EPA) have developed standard test methods to determine the toxicity and biodegradability of substances. ASTM has also developed a Guide for Assessing Biodegradability of Hydraulic Fluids (ASTM D6006) and a Classification of Hydraulic Fluids for Environmental Impact (ASTM D6046) based on the above organizations’ methods. Utilizing the methodology from these organizations, standard classifications and performance

© 2006 by Taylor & Francis Group, LLC

Lubrication in the Timber and Paper Industries

23-19

Sustainable lubricant life cycle Used oil

Seed oil

Oilcake

Oil change

Combustion

Animal feed

Seed oil processing

Oxygen

Photosynthesis Carbon dioxide

Straw

Seed oil plant

Soil

Water

Nutrients

FIGURE 23.9

Sustainable lubricant life cycle, adapted from H.F. Eichenberger [40].

TABLE 23.9 ISO Environmental Hydraulic Fluid Classifications Symbol

Classification

Commercial designation

HETG HEES HEPG HEPR

Vegetable oil types Synthetic ester types Polyglycol types Polyalphaolefin types

Vegetable oils and natural esters Polyol esters, neopentylglycols, synthetic adipate esters Polyglycols Polyalphaolefins (PAO) or synthetic hydrocarbons (SHC)

requirements for environmental fluids have also been established by the International Standards Organization (ISO). ISO environmental hydraulic fluid classifications are described in Table 23.9.

23.4.6 Vegetable Oil Based Lubricants — HETG Type HETG fluids are based on naturally occurring vegetable oils. Rapeseed (canola), soybean, and sunflower oils are the most popular vegetable oil base stocks. Vegetable oils are triglyceride esters. Like all esters, vegetable oils are composed of a molecular combination of an alcohol and carboxylic acids. In vegetable oils, glycerin is the alcohol and the carboxylic acids are genetically determined combination of saturated and unsaturated fatty acids. Many of the plants grown for oil seed production have been bred or genetically engineered for enhanced oxidation stability and low-temperature fluidity [41]. The goal of such breeding is to maximize the monounsaturated fat content of the oil seed. Polyunsaturated fats are undesirable in lubricant applications because they readily oxidize and crosslink at high temperatures. Saturated fats are undesirable because they have a high melting point. (Vegetable shortening is solid at room temperature because it contains a high percentage of saturated fats.) A comparison of fatty acids is presented in Table 23.10. The characteristics of these fatty acids directionally correspond to characteristics of their triglycerides. Since vegetable oils are composed of a mixture of triglycerides they exhibit a wide range of melting points and oxidation resistance.

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Handbook of Lubrication and Tribology TABLE 23.10 Acids

Relative Rates of Oxidation and Melting Points of Fatty

Fatty acid

Class

Chemical abbreviation

Relative oxidation rate

Melting point, 0◦ C

Stearic Oleic Linoleic Linolenic

Saturated Monounsaturated Polyunsaturated Polyunsaturated

C18 C18 : 1 C18 : 2 C18 : 3

0.6 6 64 100

69.6 14 −5 −11

Source: Adapted from Honory, L., Handbook of Hydraulic Fluid Technology, Totten, G.E., ed, Marcel Dekker, New York, p. 1133.

TABLE 23.11

Properties of Biodegradable Synthetic Esters

Ester di-2-Ethylhexylazelate di- Isotridecyladipate NPG dioleate NPG diisostearate TMP trioleate TMP triisostearate PE tetraoleate

100◦ C, cSt 2.9 5.3 6.0 7.7 9.6 12.3 12.4

Viscosity at 40◦ C, cSt −40◦ C, c P 10.3 26.7 24.3 41.2 47.5 82.6 65.6

810 21,700 — 1,500 — 4,300 —

Biodegradability, CEC L-33-A-94 >95% >90% >90% — > 90% > 90% > 90%

Source: Cognis Tech data sheet #6D-7/2000, July 2000. Tribology Data Handbook, Booser, p. 46. 1997.

HETG fluids biodegrade rapidly, exhibit excellent natural lubricity, and have a high-viscosity index. However, all natural oils contain polyunsaturated and saturated fatty acids that limit their effectiveness to an approximate range of −20 to 65◦ C. Since timber harvesting equipment can operate at temperatures well beyond these limits, HETG oils are of limited utility in woodland applications.

23.4.7 Synthetic Ester Based Lubricants — HEES Raw materials for the production of synthetic esters may be derived from natural or petrochemical sources. Because of the array of alcohols and fatty acids available for production of synthetic esters, HEES fluids have a broader range of physical properties than HETG oils. As with HETG oils, the oxidation stability of an HEES is mainly determined by the degree of saturation of the fatty-acid moiety. Saturated adipic, 2-ethylhexanoic and stearic acids are commonly used in ester synthesis. Monounsaturated oleic acid is also used in ester synthesis because torsion created by the double bond lowers the melting point and enhances low-temperature fluidity. Further enhancements in oxidation stability are made possible by selecting a polyglycol with a substituted β carbon such as trimethylolpropane, pentaerythritol, or neopentylglycol. The properties of common synthetic esters are listed in Table 23.11. HEES fluids biodegrade rapidly, exhibit excellent lubricity, possess good oxidation resistance, are low in volatility, and have a high-viscosity index. However, the ester linkage that provides a site for microbes to begin biodegradation is also vulnerable to hydrolysis. In hydrolysis, an ester is cleaved into its parent carboxylic acid and alcohol as a result of a chemical reaction with water. Since water solubility in synthetic esters is high in comparison to mineral oil, degradation due to hydrolysis is an important consideration when using these fluids [43]. For satisfactory fluid life the upper limit for in-service water content in a synthetic ester is 0.2% [44]. The performance of HEES lubricants is more than satisfactory as long as this level of water contamination is not exceeded [45].

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23.4.8 Lubrication Management Strategies Lubrication and equipment maintenance not only have an impact upon productivity in the pulp and paper industries but they are also significant factors in determining product quality. Many maintenance professionals in the pulp and paper industry have adopted concepts that are popular in Total Quality Management (TQM) programs. These concepts are sometimes described as Reliability Centered Maintenance or Total Productive Maintenance (TPM). The goal of the TPM program is to increase production while increasing employee morale and job satisfaction. TPM focuses on equipment maintenance as a strategic element part of a business’ competitiveness. Maintenance is not regarded as an inconvenience or simply an overhead. Rather, equipment maintenance is incorporated into the production schedule in order to minimize unscheduled downtime. At the same time employees on the shop floor are encouraged to take initiative in equipment lubrication, maintenance, and process improvement [46]. TPM can have a significant impact on productivity and profitability. At a large North American paper mill where downtime is valued at $10,000/h, implementation of TPM in combination with enhanced sealing technology leads to a reduction from 22 to 23 h of lost time per month to 14 h per month [47]. Employee training and management commitment are requisite elements of implementing TPM.

23.4.9 Oil Analysis Whereas employee empowerment and enhanced cooperation between production and maintenance departments are crucial to the success of TPM, oil analysis, thermography, vibration analysis, and computer assisted lubrication scheduling have become valuable technological tools. Often thermography, vibration, and oil analysis are viewed as early warning systems to predict equipment problems. While this early alert system can reduce unscheduled downtime, these tools may also be used to: • • • •

Extend oil change intervals Identify the root cause of problems Prescribe corrective action Assess the effectiveness of the remedy

One of the most effective ways to maximize oil life is to use oil analysis to scientifically determine when oil should be changed. Tests such as viscosity, infrared, and spectrochemical analysis can be used to identify the point at which oil begins to degrade, necessitating a scheduled oil change. Contamination is measured using automated laser particle counters and Karl Fisher titration. Many industrial lubricants are formulated with base oils that have been catalytically processed to remove aromatic and sulfurized chemical impurities. Removal of these compounds can double the useful life of a lubricant. The most common problem identified in oil analysis is contamination with dirt or moisture. Degradation due to oil oxidation or additive depletion is relatively rare, although sometimes degradation can be severe enough to warrant an oil change. More often than not, the lubricants used in the pulp and paper industry can be refurbished while in service through auxiliary side-stream filtration or the use of a vacuum dehydrator. Thus changing oil based upon standard time intervals can amount to an unnecessary expense.

References [1] Stenzel, G., Walbridge, T.A., and Pearce, J.K., Logging and Pulpwood Production, 2nd ed., WileyInterscience Publication, New York, 1985, chap. 6. [2] Rushton, T., Brown, S., and McGrath, T., Impact of tree-length versus short-wood harvesting systems on natural regeneration, Report FOR 2003-3, No 70, Nova Scotia Department of Natural Resources, 2003. [3] Culhane, J., personal communication, 2004. [4] Mettler, D., Tech Update: Forwarders, Logging and Sawmilling Journal, 35(3), 50–54, 2004.

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[5] Carnes, K., Offroad hydraulic fluids: beyond biodegradability, Tribology and Lubrication Technology, 60, 32, 2004. [6] Smook, G.A., Handbook of Pulp and Paper Technologists, Angus Wilde, Vancouver, 1992, chap. 9. [7] Smook, G.A., Handbook of Pulp and Paper Technologists, Angus Wilde, Vancouver, 1992, chap. 17. [8] Hink, R., Best practice lubrication for paper machines — why bad things happen to good bearings, Machinery Lubrication, 200501, 60, 2005. [9] Hink, R., Best practice lubrication for paper machines, Machinery Lubrication, 200403, 60, 2004. [10] Michael, P., What is grease? Plant Services Magazine, Volume 13, No. 10, October 1992. [11] Mistry, A., Performance of lubricating greases in the presence of water, NLGI Spokesman, 68, 8, 2005. [12] Duchowski, J.K., Messerschmitt, A.P., Needham, T., and Collins, K.G., Improvements in equipment reliability and machine performance as a result of contamination control and monitoring procedures at a major pulp and paper facility, Tribology and Lubrication Technology, 60, 56, 2004. [13] Bergling, G., Effect of insulation and lubrication on the operational reliability and temperature of rolling bearings for steam-heated papermill cylinders, Ball Bearing Journal, 233, 1, 1989. [14] Burns, B.L., Carmichael, J.D., and Bogenholm, C.A., Paper machine bearing lubrication problems and solutions — a case history, Lubrication Engineering, 33, 173, 1976. [15] Williamson, M., Options for water removal, Practicing Oil Analysis, July–August 2003, pp. 48–53. [16] Givens, W.A. and Michael, P.W., Hydraulic fluids, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 13. [17] ISO 13357, Petroleum products — determination of the filterability of lubricating oils — part 1: procedure for oils in the presence of water. [18] Day, M., Filterability testing of paper machine oils, Machinery Lubrication Magazine, November 2001. [19] Wedeven, L.D., What is EHD? Lubrication Engineering, 31, 291, 1975. [20] Dowson, D., “Elastohydrodynamics,” Proceedings of the Institution of Mechanical Engineers., 182, Part 3A, 151–167, 1967–68. [21] Snyder, D.R., Selecting rolling element bearings for modern applications, Tribology and Lubrication Technology, 60, 28, 2004. [22] Cheng, H.S., Elastohydrodynamics and failure prediction, Lubrication Science, 2, 133, 1990. [23] Jackson, A., A simple EHL film thickness equation for rolling element bearings, ASLE Transactions, 24, 147, 1980. [24] SKF Bearing Installation and Maintenance Guide, p. 65, 2001. [25] Bergling, G., Effect of insulation and lubrication on the operational reliability and temperature of rolling bearings for steam-heated papermill cylinders, Ball Bearing Journal, 233, 1, 1989. [26] Ahrens, J.F., Unpublished data, 2004. [27] Schlaefer, P.J., Unpublished data, 2005. [28] Bala, V., Gear lubricants, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 16. [29] Errichello, R., Lubrication of gears — part 4, Lubrication Engineering, 46, 231,1990. [30] Hsu, S.M., Boundary lubrication: current understanding, Tribology Letters, 3, 1, 1997. [31] Tysoe, W.T. and Kotvis, P.V., Surface chemistry of extreme-pressure lubricants additives, in Surface Modifications and Mechanism, Totten, G.E. and Liang, H., eds, Marcel Dekker, New York, 2004, chap. 10. [32] ANSI/AGMA 9005-94: Industrial Gear Lubrication, AGMA, Alexandria, VA, August 1994. [33] Moore, L.D. et al., PAO based synthetic lubricants in industrial applications, Lubrication Engineering, 59, 23, 2003. [34] Siebert, H. and Mann, U., Gear Solutions, Volume 2, No. 3, March 2004, pp. 20–27. [35] Grega, G., Making synthetics work, Lubes ‘N Greases, 10, 22, 2004. [36] Khonsari, M. and Booser, E.R., New lubes last longer, Compoundings, 55, 21, 2005.

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Lubrication in the Timber and Paper Industries

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[37] Michael, P.W., Herzog, S.N., and Marougy, T.E., Fluid viscosity selection criteria for hydraulic pumps and motors, Proceedings of the 48th National Conference on Fluid Power, NFPA, Milwaukee, 2000, 313. [38] Placek, D., Study examines multigrade fluids for forestry equipment, Hydraulics and Pneumatics, 54(3), 39, 2001. [39] Herzog, S.N. et al., Effect of operation time on oil viscosity and pump efficiency, NCFP I05-9.3, IFPE March, 2005, Las Vegas, NV, USA. [40] Eichenberger, H.F., Biodegradable hydraulic lubricant: an overview of current developments in central Europe, 42nd Earthmoving Industry Conference, Peoria, SAE 910962, 1991. [41] Maelor Davies, H. and Flider, F., Designer oils, Chemtech, April 1994, pp. 33–37. [42] Buenemann, T.F. et al., Non-aqueous synthetic lubricants, in Fuels and Lubricants Handbook, Totten, G.E., Westbrook, S.R., and Shah, R.J., eds, ASTM, West Conshohocken, PA, 2003, chap. 10. [43] Kempermann, C. and Murrenhoff, H., Reduction of Water Content in Biodegradable and Other Hydraulic Fluids, 1998 Earthmoving Conference, Peoria, SAE 981497, 1998. [44] Gere, R.A. and Hazelton, R.A., Polyol ester fluids, in Handbook of Hydraulic Fluid Technology, Totten, G.E., ed., Marcel Dekker, New York, 2000, chap. 19. [45] Young, K.J. et al., Environmental standards for biodegradable hydraulic fluids and correlation of laboratory and field performance, Lubricants for Off-Highway Applications, SAE SP-1553 (2000-012543), pp. 15–21. [46] Venkatesh, J., An introduction to total productive maintenance, Plant Engineering, February 2005. [47] Nooyen, J.G., personal communication, 2004.

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24 Textile Fibers/Fabrics 24.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24.2 Friction and Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

24-1 24-2

Fundamental Concepts • Testing Methods • Experimental Observations

24.3 Textile Processing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

24-11

Spinning Process • Texturing Process • Knitting Process • Weaving Process

24.4 Fiber Finishes and Components . . . . . . . . . . . . . . . . . . . . . .

Paul D. Seemuth Tribology Consulting International LLC

24-16

Formulated Finish Package • Lubricants • Surfactants (Emulsifiers) • Antistats • Other Finish Components • Coning Oils

Acknowledgments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

24-26 24-27

24.1 Introduction The introduction of synthetic polymers has brought forth a wide range of applications to the modern world. Along with the various uses of these systems, tribology, the science of friction and wear, has had a pronounced effect on the knowledge of the properties achieved and the end-use possibilities for these materials. A major application of synthetic polymers influencing the world’s everyday life is textile fibers and related textile constructions. Textile materials are made from a variety of polymer types. The widely known textile related polymers are polyesters, polyamides, polyolefins, polyurethanes, and aramids. Fiber manufacturers produce a variety of chemical compositional forms within each class and introduce unique additives for special use characteristics. One fundamental challenge faces all fiber producers. These materials experience physical contact with numerous surfaces as well as rub against themselves. In contact with other fibers, other metal, or plastic surfaces, resistance to movement across the surface is noted and defined as friction. In a typical processing step, the number of surfaces and their nature, metal, hot, cold, smooth, rough, act as a destabilizing combination of friction forces. Therefore, the frictional forces imposed on these soft “plastic-type” materials frequently result in a number of product defects. To produce a quality product, friction needs controlling in a way to give uniform stress loads during use. A uniform and balanced friction load during processing eliminates most defects, that is, broken filaments, nonuniform fiber structure, static, and finally package formation. Unlike their naturally lubricated fiber counterparts, wool, cotton etc., the synthetic fibers require lubricants to control friction that is crucial to a formation of a quality fiber and its subsequent use. Therefore, application and extension of the principles learned in other business, that is, natural fiber and metalworking, were crucial for these new materials. 24-1

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Handbook of Lubrication and Tribology

During the early 1900s, textile lubrication, fiber finishes that are combinations of a multitude of different chemical components, became key to the successful commercialization of the new textile fiber business. However, the unique character of synthetic polymers needed more than the old lubrication fluids of the past. Maintaining good yarn color, meeting processing demands, and imparting a soft nature to the fibers themselves demand new lubricants and auxiliaries that are suited to the production conditions and final use of these fibers. To overcome many processing issues, fiber finishes are normally applied as water-based emulsions. These emulsions can be used at concentrations from 0.1% active to >50% (typically 5–20%), dependent on the fiber type, process, and processing speed. Typical loading on a fiber is from 0.05% (staple) to 3–5% for polyurethane systems. Further, the finish must combine a number of physical properties, good “uniform” wetting on fiber surface and associated fiber bundle (wicking), low aerosol generation (environmental, health, and safety), static dissipation character, and finally good overall balanced frictional properties. These parameters are only a small set of the physical characteristics that the fiber finish must incorporate for use on the fibers. Improper cohesive forces may result in poor staple yarn spinning. It may result in a package of yarn being too hard or soft resulting in performance problems during knitting or weaving. “Sluffing,” multiple package yarn layers being removed at the same time, and tension plucks, high friction forces between the fibers themselves or filaments trapped underneath each other encountered during yarn removal from package, are common quality problems associated to poorly formulated finish systems. The list of finish characteristics is long and complex and many features are discussed later. Today’s concept of the finish must be an integral system component of the product. One will discover during work on fibers that the surface lubricant system is just as crucial as the underlying “solid” substrate for the final product, be it a garment to wear or a life saving bullet proof vest.

24.2 Friction and Lubrication 24.2.1 Fundamental Concepts Friction is defined as the resistance of movement of one body moving against another. This concept covers many contributing features of friction, apparent contact area, relative speed, sliding, and rolling cases. From Leonardo da Vinci’s work (1452–1519), the first law of friction evolved three key points: 1. Friction is independent of the area of the solids. 2. Friction is directly proportional to the load that is normal in the direction of movement. 3. Frictional force is independent of the speed. Therefore, the first law is expressed in Equation (24.1) as F = µW

where F = Friction and W = Load

(24.1)

This amounts to the coefficient of friction, µ being a constant, independent of load (W ). This general assumption works for most situations in first approximations though the frictional properties exhibited by real dynamic systems show variations from these simple laws. Work in the later 1800s and 1900s refined the knowledge of the frictional forces that are in play during the contact of two surfaces, normally one stationary relative to the second body. Second and Third Laws of Friction have been developed and studied and reviewed extensively [1]. Various authors [2–8] have described fiber friction. This frictional behavior is illustrated by Stribeck’s curve shown in Figure 24.1. The forces experienced by a fiber fall into four different regions. Of primary importance to most textile processes are the boundary and hydrodynamic regions. Both of these regions affect most of the properties of the yarn though the extent is variable over a considerable range. Examining Figure 24.1, boundary lubrication (boundary frictional forces) diminishes as the speed of the yarn accelerates. Similarly, hydrodynamic frictional forces increase though they do reach a temporary stable value plateau

© 2006 by Taylor & Francis Group, LLC

Textile Fibers/Fabrics

24-3 Basic friction regions

Boundary region

Coefficient of friction

Semi-boundary region

Hydrodynamic region

Solid-like region

[(Thread line speed) (Lub viscosity)] Pressure

FIGURE 24.1 Stribeck’s curve — general representation of frictional behavior of lubricated textile yarns (updated from Reference 21, p. 5, courtesy of Marcel Dekker, Inc. and Reference 17, p. 117, courtesy of Kao Corporation).

Low inter-molecular shear forces allowing lubricant to shear with itself prior to interfacial shear

Absorption of boundary lubricant to substrate by intermolecular forces (Van der Waals)

FIGURE 24.2 Representation of sheared lubricant layer under boundary conditions.

as the speed increases. The boundary and hydrodynamic region overlap at a minimum and this “semiboundary region,” a combination of both regions, operates in ill-defined manners. The main two regions are characterized and actually controlled independently using proper component selection by a finish formulator. Boundary lubrication (Fy−y or Fy−m ) is managed by application of an intramolecularly low shear material on the fiber surface [9–12]. This low shear material has a reasonable high affinity to stay “fixed” onto the polymer surface and it functions by shearing layers of itself from one another rather than shearing from the surface. Figure 24.2 best illustrates this. Under boundary conditions, the textile polymer is in close intimate contact with the other surface, be it another filament or an external contact point. During movement of the fiber, the lubricant film adheres to the fiber, the lubricant itself shears intramolecularly which accounts for low friction forces under highpressure and low speed conditions. Furthermore, this effect is necessary for damage prevention of the delicate fiber surface and uniform distribution of the stress load across each filament in a fiber bundle. The semi-boundary regions (stage two) are the less understood of the friction regions. Here, the relative speed change of the two surfaces introduces multiple interactions of both boundary and dynamic friction parameters and reaches a minimum value at this stage [2]. The third stage is the dynamic region (Fh ). It is here that textile processing requires considerable attention. The two bodies are in a high degree of relative motion, one normally stationary to the moving

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Handbook of Lubrication and Tribology

Low shear strength at fiber–lubricant interface allowing fiber to ride (surf) on the lubricant “pool”

FIGURE 24.3

Representation of sheared lubricant layer under hydrodynamic conditions.

body, or fiber. This region is highly sensitive to lubricant viscosities, its inherent resistance to shear [13,14]. This is illustrated in Figure 24.3. In this region, the fiber glides on a “pool” of lubricant across the other surface. The shear point is at the fiber/lubricant interface, as the fiber/lubricant attractive forces are minimal. For the fiber finish formulator, the combinations of components serving multiple functions must be factored together to understand the final dynamic frictional values obtained. In the fourth region, hydrodynamic friction at very high speeds, work by Kao Corporation [17] in Japan suggests that the frictional forces start to decrease in what is termed the solid-like region. For most new textile processes, these domain speeds are now being explored further. Several hypotheses are presented to account for the decrease in friction. Friction decreases with speed for these possible reasons: 1. Speed results in slippage between the lubricant layer and the friction contact vs. lubricant shear. 2. Dynamic mechanical properties of the lubricant vs. the steady-state lubricant viscosity controls friction [14]. 3. High temperatures produced at extreme speeds further reduce the apparent lubricant viscosities, a controller of frictional forces.

24.2.2 Testing Methods Evaluation of frictional properties continues to receive considerable attention. The major work of the 1950s to 1990s focused on the macro scale for friction measurements [2–15]. The past 15 years has brought more focus on the microscopic as well as the nanoscale properties of frictional behavior [19]. All these areas are fundamentally crucial to the understanding of behavior; however the textile industry relies heavily on the macroscale effects. These effects are readily measurable and relate well to the daily processing behaviors. With the dependence on the practical aspects of friction for everyday processing, it is crucial to understand the available tools and testing protocols allowing rapid and consistent testing of the lubricants, surfactants, and other additives that are addressed later in this chapter. Evaluation of the fiber’s frictional performance containing a lubrication package requires the finish application onto a substrate of interest. Lawson-Hemphill provides a commercial yarn re-winding unit, a Precision Lab Winder, and the simple inclusion of a finish pump or syringe pump allows an effective adaptation for accurate finish application to a substrate. The best choice of substrate is specifically the polymer material that is under examination, be it polyester, polyamide, aramid, polyolefin, or some other type. There is always the dilemma for the formulator to have access to a suitable material containing no surface lubricants or other add-ons. However, frictional assessment within a fiber substrate set using a working process control provides an excellent relationship to another substrate set. Studies have been done which

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Textile Fibers/Fabrics

24-5

0.7 0.6 0.5

Fh

0.4 0.3 0.2 0.1 0 PET

Nylon

Cellulose acetate

Poly acrylonitrile

None

Hexadecane

Butylstearate

Oleic acid

Ethyl stearate

Octadeyl alcohol

Stearic acid

Octadecylamine

Solid stearic acid

Solid octadecylamine

FIGURE 24.4 Frictional response of different polymer surfaces (adapted from Reference 10, p. 1143, American Chemical Society).

illustrate this point. Schick studied the effect of substrate on friction. Though he was examining other attributes and these must be taken into account for the relationship to be binding, the effect is translatable [9]. The study shows that each polymer substrate has a different frictional response (Figure 24.4). Considering this, a single available control fiber material, using a control lubricant package, allows formulation of the systems to meet effectively the demands to a polymer type and its processing. Boundary (Fy−y ) and hydrodynamic (Fy−m or Fh ) friction are evaluated a number of ways. Commercial equipment for running frictional analyses can be obtained that is highly accurate with good precision. During the past years, many of the original frictometers have been discontinued or updated with modern computer data analysis features. Rothschild, Zurich Switzerland, markets the more notable of the newer instruments, upgraded from its long-term line of instruments devoted to the textile industry. LawsonHemphill Inc. provides a new design frictometer that allows improved pre-tension control during frictional testing. Both frictometer systems offer an excellent speed range for hydrodynamic (Fh ) evaluation up to 1000 m/m or 550 m/m respectively (Figure 24.5 and Figure 24.6). A schematic drawing of the Rothschild unit is shown in Figure 24.7. Friction is calculated from the input tension T1 and the resulting output tension T2 . These values give the coefficient of friction using the following Equation (24.2) [2,20–22]. T2 /T1 = eθ µ

(24.2)

Written in a more useful form, the coefficient of friction is expressed in Equation (24.3): µ=

ln(T2 /T1 ) θ

(24.3)

where θ is the wrap angle expressed in radians [3,23,24]. Graphical representations of the friction coefficient are best expressed by plotting the output tension, T2 , against increasing speed. One important reason for this is the calculation for µ will give the same

© 2006 by Taylor & Francis Group, LLC

24-6

Handbook of Lubrication and Tribology

FIGURE 24.5

Rothschild frictometer R-2088 (photo courtesy of Rothschild Instruments, Zurich, Switzerland).

FIGURE 24.6

Lawson-Hemphill constant tension transport (photo courtesy of Lawson-Hemphill, Inc.).

number over a range of T2 values (at a given T1 input level). This is especially true under high T1 inputs. The false range of calculated µ masks the real frictional differences of experimental lubricant systems, which, if one translates the in-lab data to an operational process, will lead to inconsistencies in the results expected.

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Textile Fibers/Fabrics

24-7 T1 input tension head

T2 output tension head

u u expressed as radians for 1708

u u expressed as radians for 808

FIGURE 24.7

Thread line diagram Rothschild frictometer R-2088.

Using these calculations and graphical representations, friction performance can be evaluated for a set of lubricant systems. Additionally, the Rothschild and Lawson-Hemphill equipment allows modified setups so that different friction surfaces, process temperature environments, and contact angles are available in the thread path. The friction pin surface can be readily changed to a variety of chrome, ceramic, and other surface types along with modifications of surface roughness that will give even better experimental basis for textile operations and associated process behaviors. This provides detailed analyses of a finish lubricant’s properties under a variety of processing conditions. These frictometers have the capability of low speed operations for analysis of the boundary friction region. However, most of the boundary friction test methods, usually a three-twist configuration to the friction test surface, are limited to the types of yarns being studied [25–28]. POY (Partially Oriented Yarns) will undergo stretching during testing giving variability in friction values. FDY (Fully Drawn Yarns) provide the best overall stability under this three-twist configuration. In the early 1960s, a highly reliable method was developed by workers at DuPont [2,12]. With many improvements over the years, the primary equipment for this testing still is custom-made due to the nature of the testing, the importance placed on understanding this friction region, and the unique friction relationships crucial to the process. A typical unit is shown in Figure 24.8. The Capstan method uses a rotating mandrel containing a wound tube of yarn to move at speed ranges of from 0.0016 to 32 cm/sec. Over the mandrel is draped a single strand of fiber (from the same yarn/lubricant test source) and is attached to a strain gauge and held under constant T1 . The mandrel rotates and the frictional forces are recorded on a strip chart recorder or are recorded digitally. Using this method, the phenomenon of stick slip is easily observed. Stick slip is a characteristic of the overlapping nature of boundary, semi-boundary and hydrodynamic lubrication regions under slow speed conditions as seen in Figure 24.1. The effect that arises is a combination of several chemical and physical factors: 1. Shear strength of the fluid film 2. Adhesion forces between fibers (at dry contact regions resultant for nonuniformities in film thickness) 3. Van der Waals interactions (any interactions between uncharged molecules) These effects diminish as these forces are overcome and completely dampened with increasing speed. Figure 24.9 illustrates this. A good balance of the stick-slip friction magnitude will enhance the ability of the yarn to form a good quality yarn package and assist in uniform delivery of yarn off the package.

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24-8

Capstan frictometer (photo courtesy of INVISTA, Inc.).

Fm

FIGURE 24.8

Handbook of Lubrication and Tribology

10–3

10–1

10 Increasing speed

102

FIGURE 24.9 Typical stick-slip phenomena representation (reprinted from Reference 7, p. 105, courtesy of Textile Research Journal).

In the evaluation of yarns, there are other parameters that one needs to access. Static is a principle factor, along with friction, to control in many processes. Finish systems will have varying degrees of effectiveness to dissipate static charge that is built-up as fiber runs over guide surfaces and rubs against itself. Various methods have been outlined for assessing static charge dissipation with a fiber finish. One method was outlined by Schick [29]. The yarns are wound onto skeins and a voltage of 160 V was applied. The time is then measured for a drop to half value of applied voltage. Other methods have been applied for measurement of antistatic properties. Usually these results will be given as resistivity or as log Rp . In Figure 24.10, the general region for good static control in knitting and staple process is below log Rp of 9. A mid range of 9 to 11 will handle processes such as texturing. Once the value starts to rise above 11, static generation problems become a severe concern. While this serves as a guideline, measurement using the induced voltage charge and relaxation time to 12 values is an excellent basis. Under hydrodynamic

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Textile Fibers/Fabrics

24-9 14 13 12

Log Rp

11 10 9 8 7 6 5 excellent

good

fair

poor

Degree of antistat protection

FIGURE 24.10

General relationship of log R vs. static protection performance.

conditions, the Rothschild R-2088 unit has the capability to measure static charge generated at the friction pin surface using an electrostatic voltmeter. The Lawson-Hemphill CTT units have also adapted the frictometer in various ways to measure the electrical charge generated at the guide surface. As the charges are small, typically pico-ohms, consistent measurements require that care be taken to control the equipment’s environment and conditions of testing. Running the yarn at various speeds gives an accurate picture of the finish’s ability to dissipate charge under hydrodynamic conditions. A well-defined relationship between hydrodynamic friction, charge dissipation, and finish properties is achieved by collecting data from 50 to 1000 m/m.

24.2.3 Experimental Observations Numerous parameters influence the frictional response of a fiber bundle as it is moved over guide surfaces and the frictional effects are noted from the time the polymer exits the spinneret for fiber formation until the final fabric or garment is packaged for use. Several of these have been covered — speed of the yarn during process, process temperature, contact area between yarn and guide, tension or pressure at contact point which is a function of the contact angle influenced by guide diameter. Others include denier of fiber, finish viscosity and related physical properties, yarn luster (different surface roughness due to delusterant inclusion), and guide roughness [7,15,16]. These contributors play major roles in the overall requirements for lubrication balance during processing of synthetic fibers. Others that contribute to processibility are relative humidity and environmental temperatures that affect static propensity, moisture regain, and plasticization of a fiber’s surface (especially with nylon) [30–33]. Surface roughness is a major outside influence that affects the frictional forces experienced by a fiber [7]. While the focus is normally directed to the fiber system, that is, polymer and lubricant (finish), the roughness of the guide surface has as much influence as the changing luster of the fiber itself. As noted earlier, for a selected group of polymer types, PET friction was lowest of the group studied (see Figure 24.4). In the same way, within a polymer type, the hydrodynamic frictional forces are lowered for a fully dull yarn vs. a bright yarn. This is due to the decrease in surface contact and proposed theory on disruption of the fluid film between the guide and the fiber [1,4,6]. Reversing the situation by changing the roughness of the guide gives a similar effect as illustrated in Figure 24.11 [2,12,18]. With a given polymer type, increase guide roughness decreases the coefficient of friction quite dramatically. A drop of 10 to 60+% can be seen. When this is translated into a textile operation, guide roughness or multiple guides of different roughness along the running thread path can alter the fiber’s performance. It is in this world of nonuniform guides that the balance of frictional forces through lubrication is most demanding.

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24-10

Handbook of Lubrication and Tribology 4.6 micro inch Chrome

Coefficient of friction

8 micro inch Chrome 40 micro inch Chrome 40 micro inch Alsimag

90 micro inch nodular Chrome

Speed in yards/min

FIGURE 24.11 Frictional effect of guide roughness modification (data courtesy of INVISTA Inc.). (From Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969).

Viscosity cps at 258C

120 100 80 60 40 20 0 0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

Coefficient of friction

FIGURE 24.12 Relationship of viscosity to hydrodynamic friction for 25 commercial fiber finishes (Seemuth unpublished data, courtesy of INVISTA Inc.).

Viscosity of the lubricant package is the other major influence on frictional forces [13]. As noted in Figure 24.1, friction forces are directly relatable to the viscosity of a finish. While there are extenuating cases, first principles and numerous studies have shown a good relationship of hydrodynamic friction to viscosity measured using a Brookfield viscometer. A good example of this relationship is shown in Figure 24.12. A study of the friction vs. FOY on 40-13 dull nylon further illustrates the viscosity effect. Friction starts to reach a plateau at FOY levels of 0.8 to 1.0% (Figure 24.13). This is attributable to the increased shear stress of the total fluid film volume separating the surfaces, that is, a leveling out of the shear forces where lubricant‘s viscosity reaches an equilibrium value [13,20,22]. Once the plateau is reached, additional lubricant levels afford no additional value. One other premise in this explanation of the frictional curve profile is that the wetting of the lubricant on the surface forms a continuous fluid film in accord with the thermodynamics of wetting [34–37]. If the film was nonuniform, the friction forces would be governed by fiber surface properties and discontinuities of the fluid’s film. With modern textile operation reaching speeds rapidly approaching 6000 m/m, effects of temperature play an important consideration in the balance of lubrication properties. Viscosity decreases as the temperature increases. While there are limits to the viscosity value reached, temperature of the process can alter

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Textile Fibers/Fabrics

24-11 0.8

0.75 0.7 0.65 Fh

0.6 0.55 0.5 0.45 0.4 0.35 0.3 0.1

0.24

0.37

0.47

0.69

1

1.55

2

% FOY

FIGURE 24.13 Hydrodynamic friction vs. FOY on 40-13 dull nylon (data courtesy of INVISTA Inc.). (From Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969.)

the performance. This temperature effect is also dependent on the molecular weight and thermal stability of the lubricant. Increasing temperature using low molecular weight lubrication of low thermal stability will result in changing composition and lubrication character as the lubricant volatilizes and decomposes. With the high speeds and increasing temperatures of processing, thermally stable lubricants are finding increasing use for textiles. These stable lubricants lower the rate of change of the compositional make up of the lubricant package and stabilize the frictional forces on the yarn during process transient. This effect is the result on less percentage loss of the lubricant, less decomposition of lubricant, and retention of the associated frictional properties. It is evident that the friction regions, especially boundary and hydrodynamic, are distinct for a synthetic fiber running against itself or an external surface. It should be further pointed out that while there is overlap of the regions, detailed work on the primary regions, boundary and hydrodynamic, provide most of the crucial information for controlling the friction balance necessary for a textile fiber to perform to it’s full capability during processing. While various references and researchers [2,23,29] provide slightly diverse ranges for the different regions, a general guide is presented. Boundary friction is attributed to speeds under 0.1 m/min. The semi-boundary region range is from 0.1 to 5.5–10 m/min. The hydrodynamic region is speeds exceeding the 5.5–10 m/min. While I do not understate or underestimate the importance of the underlying polymer structure or the external guides and processing conditions, textile lubrication researchers must realize that the fiber lubricant is an integral part of a widely interactive system, dependent on many parameters and independent on almost none. Therefore, the lubrication properties must match the process, be compatible with the substrate, and exhibit all the correct physical attributes necessary for optimum processibility [38]. Examining the delicate balance needed for the lubricant system, several key parameters surface that influence the overall lubrication quality of the fiber product. Dominating the boundary lubrication region are wetting and geometry of the surfaces in contact. Viscosity, wetting, and thermal stability are crucial parameters for hydrodynamic lubrication at high speeds.

24.3 Textile Processing 24.3.1 Spinning Process Polymer systems are produced by three major routes; melt, dry, and wet spinning [39–42]. All these routes are extensively covered in literature and will only be briefly mentioned for completeness. Thermoplastic polymers are normally melt spun and produced as multifilament bundles. Most nylons, polyester, polylactides [43–45] and polyolefins are pre-polymerized, then formed in chips used for later re-melt

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spinning or using a continuous process are produced directly into the filament form. Dry spinning is almost a misnomer as the polymer is dissolved in a solvent which flash evaporates upon exiting the spinneret. The polyurethanes fibers, Globespan® [46] and Lycra® [47] spandex fibers are examples of this type of process for fiber formation. Wet spinning is the most complex of the spinning techniques. Some wet spinning process requires the polymer solution to be kept above or below ambient temperature and thus a heat exchanger is used. The spinnerets are just above or immersed in tanks containing the coagulation medium into which the filaments are extruded. Wet spinning spinnerets can have up to 2000 holes. Wet-spun polymer types that are available include aramids, like Kevlar® [48] and Nomex® [48], viscose and acrylics. To produce the initial continuous fiber bundle, the polymer mix is pushed through a multi-holed spinneret having unique hole designs. Routinely, filaments are produced in round, trilobal or octalobal forms. Along with this common commodity set, modern spinneret cutting techniques can make a seemingly endless variety of shapes and filament sizes in recent years. Into the mid 1980s, deniers of >1.5 dpf were the most commonly produced. With the advent of more critical consumer demands, unique fibers were born, such as Nylon Supplex® [47], Coolmax® [47], MicroMattique® [47], Polarguard® [49], and Meryl® [50] Microfibre and Nateo. These are 5 carbon units as represented in the Ucon® and Pluronic® systems, and finally the anionic or cationic surfactants.

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Surfactancy

High

R=Alcohol

R9 = Acid

Low Increasing moles of ethylene oxide

FIGURE 24.22 oxide (EO).

General trend of surfactancy effectiveness of ethoxylated alcohols and acids vs. moles of ethylene

Each of these systems adds unique properties to the finishes. These materials may function as the primary emulsifier, co-emulsifier, or coupling agent. The first surfactant class is the nonionics [74–76]. These systems are prepared from an acid or alcohol unit, the hydrophobic part, and ethylene oxide forming the hydrophilic unit. These nonionic surfactants are widely used as the primary or the secondary (coupling) emulsifier systems for a fiber finish though their effectiveness does vary based on structure, hydrophobe [77] and EO unit length. Figure 24.22 illustrates the effectiveness of the acid vs. alcohol initiated ethoxylated materials. As the number of moles of EO increases relative to the molecular weight of the hydrophobic group, the ability to emulsify the hydrophobic oil also increases. In this regard, there is an EO content level where the surfactant transitions from a liquid to a solid and the ability to formulate with these materials may be hindered by oil and surfactant solubility issues. In these cases, the use of several compatible surfactants, one functioning as a so called co-emulsifier, will be necessary to provide a stable oil formulation as well as being able to give stable emulsions of the primary lubricant and other materials. In most cases, the molar range of EO finding most use for the mono alcohols or acids is 5 to 15 EO units. This is common for alcohol or acids of 6 to 20 carbon atoms. For polyhydritic alcohols, meaning having more than one alcohol moiety, the range of EO units will range from 0 to 200 EO units. This class of surfactants can best be represented by R–(C)a (O)x -(EO)y –Rz where R is hydrogen or alkyl, a is 3 to 6, x is 2 to 6, y is 0 to 200, R is H or alkyl, and z is equal to (x − 1). Several commonly used surfactants of this class are the glycerol monoester, glycerol mono-oleate (GMO) or glycerol mono-isostearate (GMiSt) and ethoxylated sorbitol mono or the multi-esterified esters. These nonionic surfactants from the multifunctional alcohol class give improvements in thermal stability due to their higher molecular weights, have a wider liquid state range per EO content than the simple straight chain emulsifiers, and exhibit good compatibility with most lubricants. These materials are widely available from many suppliers in many variations for the finish formulator. The second emulsifier class is the alcohol or acid PO–EO systems [78,79]. The availability of these is limited and used for special processing cases. While offering decent emulsification properties, the ratio of EO and PO will be the key parameter as to the extent of utility without excessive use of co-emulsifiers. Figure 24.23 represents the typical effect of emulsification potential for the various structural possibilities. The blocked systems with the EO unit at the end are the best surfactants of the series. This arrangement of structure allows the hydrophobic alkyl and PO units to align with the hydrophobic lubricants and the hydrophilic EO units to the corresponding hydrophilic materials, the other emulsifiers for oil compatibility or water at the surface of the micellar emulsion structure. One key property that these materials can provide is the lowering of chemical absorption (chemical migration) into the soft polyurethane or rubber machine parts. Whereas the simple ethoxylated materials have a high absorptive character to the typical PU and rubber machine parts, substitution of these EO–PO surfactants, though slightly lessening the overall surfactant package emulsifying potential, greatly lowers the damage related to the softening

© 2006 by Taylor & Francis Group, LLC

24-23

Emulsification potential

Textile Fibers/Fabrics

-EO

PO R-

O

O-P

R-E

om

Rand

% ethylene oxide [fixed PO content]

FIGURE 24.23 General trend of surfactancy effectiveness for alcohol or acid EO–PO systems.

360

Visocsity in cps

340 320 300 280 260 240 220 200 0

1.5

3

5

10

Moles of PO for capping

FIGURE 24.24 PO end-capping viscosity effect on blocked alcohol PO–EO surfactants (data courtesy of Milliken and Co.).

of the PU and rubber machine parts, abrasion, wear, and altered physical properties. As can be expected, the increased molecular weights of these types of materials will raise the overall viscosity of the finish package and consequently, the frictional properties on the polymer surface. An additional enhancement available to the finish formulator is the ability to modify the structure of these systems altering the viscosity of the individual materials. While this route has several limitations, the effect can be dramatic and useful when improved compatibility with oil formulation and machine parts is a critical parameter to success. This effect is illustrated in Figure 24.24 as well as the limitation. Taking a base alcohol PO–EO system, simple capping with an additional PO molar ratio of 1.5, a 20 to 50 cSt unit or more dropped in viscosity can be obtained. However, this effect is quickly lost with increased PO addition. While this example is larger than most, it is a valuable synthetic modification tool to assist in maintaining frictional properties at improved compatibility with other materials. A similar effect also appears with other simple ethoxylated materials and the multifunctional alcohols. Furthermore, the PO unit preserves the emulsification and frictional properties of the related and widely used MPEG esters. These PO capped esters exhibit none of the incompatibility with certain polymer types nor exhibit some of the topical numbing effects known for many methyl capped EO alcohol systems. The last classes of surfactants are the anionic [80–82] and cationic [83–85] materials. Again, a wide variety of materials is available and several will be mentioned in the discussion on antistatic components.

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These materials were the first of the classes of surfactants used by finish formulators. Many were vegetable oil based and while use has decreased with time, they still form an integral part of the formulator’s chemical toolkit. Several of the most common are the sulfated peanut and vegetable oils. These provide good compatibility with various mineral oils and continue to be used in many cases. As process speeds have increased along with needs for low finish oil color, these materials are slowly being replaced in favor of the aforementioned nonionic surfactants where necessary. One widely used anionic surfactant is sodium di-octylsulfosuccinate, commonly referred to as DOSS. This material is extensively studied for its surfactant and wetting properties [86]. In many finish systems, the presence of DOSS allows for improvements in wetting, lower emulsion, and oil surface tensions, without the need of the expensive fluorochemicals or silicones. With its moderately high molecular weight, thermal stability, anionic stabilization of emulsion micelles, and low cost, this material continues as a mainstay surfactant for finishes.

24.4.4 Antistats Antistats incorporated into a finish systems manage and minimize electrical charge generation during processing. These materials introduce hygroscopicity along with the surfactants to the lubricant system. With these materials present, the nonconductive polymer surface possesses the functionality allowing discharge of generated electrical static into the surrounding environment, mainly into the air’s moisture. Both the surfactants and the antistats contribute to the overall antistatic properties of a fiber. The typical surfactants are able to assist in charge removal from the polymer’s surface though their function relies solely on the lone electron pairs of the oxygen atoms of the EO units. As these materials have limited dissipation propensity, many formulations, especially those used for knitting and weaving, incorporate the stronger anionic or cationic classes of antistats. These highly charged materials, provide rapid and effective elimination of static charge. Typical classes of materials are shown in Figure 24.25. Of these, the phosphates and quaternary amines comprise the most widely utilized materials. These two types of chemical systems allow good compatibility in most finish packages. Most widely used are the phosphates with the ammonium salts being used more for processing conditions where the control of processing humidity is less likely, that is, older plants or global facilities in undeveloped countries. Formulation with these materials is a challenge. Dependent on the nature and size of the R groups, solubility in primary finish lubricants can be tricky. Most ammonium systems formulate well though their overall effectiveness is lower than the corresponding phosphates. The phosphates, on the other hand, are most effective, not as the free acid, but as the corresponding salt though oil compatibility becomes an issue. The potassium cation is the preferred counter ion for the phosphate anion. The cation–anion pair provides a larger sized ionic sphere associated with improved static elimination. When the sodium salt, a small and closely associated ion pair with anions, is tested against the corresponding K+ salt, the latter is generally more effective to dissipate charge. Furthermore, as the MW of either system increases, the charge dissipation effectiveness decreases. This is the effect of charge density vs. molecular size. Using many of the very hydrophobic lubricants, the larger R groups afford better oil compatibility than the smaller phosphate salts. The R group choice is also dependent on the fiber type and structure. Low orientation, high elongation fibers will absorb the lower molecular weight though more highly effective antistats. • Anionic R2P(O)O– RCO2– RS+O3–

• Cationic R4N+X–

• Amphoteric • Nonionics RO-(CH2CH2O)x-H R-(PO)x-(EO)y-H

FIGURE 24.25

Chemical classes used for textile antistats.

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R3N+CH2CO2–

Textile Fibers/Fabrics

24-25

13 12

Log Rp

11 10 9 8 7 6 250 280 325 355 365 380 390 410 415 455 490 530 Antistat molecular weight Phosphate at 22% Rh

Quat Ammonium at 22% Rh

Phosphate at 47% Rh

FIGURE 24.26 Effect of relative humidity on antistat effectiveness (Seemuth unpublished data, courtesy of INVISTA Inc.).

With Low Orientation Yarns (LOY) and many nylon staple products, phosphates based on alkyl groups 170◦ C) and sometime very high temperatures (>300◦ C) volatilizes off much of the finish systems, forcing applications of additional lubricants to facilitate the next processing stage.

Acknowledgments I wish to acknowledge E.I. du Pont de Nemours and Co., Inc. and INVISTA Inc. for providing me the opportunity to work in the synthetic fiber field and develop the science of Fiber Tribology. I further wish to acknowledge several mentors who have guided my study and work in the field of tribology, polymer and finish sciences, and toxicology especially Brian Briscoe, Imperial College, London; Karl Jacobs, Georgia Institute of Technology; Euan McClelland, M. Godwin Jones and Robert Phillips, DuPont retirees; and Gerald Kennedy and John Gannon, DuPont. I gratefully thank my other colleagues around the world and all the contributors to this chapter.

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References [1] Hutchings, I.M., Tribology: Friction and Wear of Engineering Materials, Edward Arnold, London, 1992, p. 23. [2] Olsen, J.S., Frictional behavior of textile yarns, Textile Res. J., 39, 31, 1969. [3] Howell, H.G., Meiszkis, K.W., and Tabor, D., Friction in Textiles, Butterworths, London, 1959. [4] Cameron, A., Basic Lubrication Theory, Ellis Horwood, London, 1981. [5] Baird, M.E. and Meiszkis, K.W., J. Textile Inst., 46, 112, 1955. [6] Schick, M.J., Friction and lubrication of synthetic fibers, in Surface Characteristics of Fibers and Textiles, Schick, M.J. (Ed.), Marcel Dekker, New York, 1975, pp. 2, 3, 5. [7] Schick, M.J., Friction and lubrication of synthetic fibers part I: effect of guide surface roughness and speed on fiber friction, Textile Res. J., 43, 193, 1973. [8] Schick, M.J., Friction and lubrication of synthetic fibers part II: two component systems, Textile Res. J., 43, 198, 1973. [9] Bowden, F.P. and Tabor, D., The Friction and Lubrication of Solids, Clarendon Press, Oxford, Part 1, 1950 and Part II, 1964. [10] Ford, Jr., T., Adsorption and boundary friction on polymer surfaces, J. Phys. Chem., 66, 1136, 1962. [11] Ford, Jr., T., Adsorption and boundary friction on polymer surfaces, J. Phys. Chem., 44, 1136, 1962. [12] Ford, Jr., T., and Olsen, J.S., Boundary friction of textile yarns, Textile Res. J., 31, 1007, 1961. [13] Schick, M.J., Friction and lubrication of synthetic fibers Part IV: effect of fiber material and lubricant viscosity and concentration, Textile Res. J., 43, 342, 1973. [14] Park, H., Seefried, Jr., C.G., and Bryant, G.M., Relation of lubricant structure to frictional properties — polyoxyalkylene monoether lubricants on filament yarns, Textile Res. J., 44, 692, 1974. [15] Schick, M.J., Friction and lubrication of synthetic fibers Part V: effect of fiber luster, guide material, charge and critical surface tension of fibers on fiber friction, Textile Res. J., 44, 758, 1974. [16] Schick, M.J., Friction and lubrication of synthetic fibers Part III: effect of guide temperature, loop size, pretension, denier, and moisture regain on fiber friction, Textile Res. J., 43, 254, 1973. [17] Kao Corporation, Surfactants — A Comprehensive Guide, Tokyo, 1983, p. 107. [18] Slade, P.E., Handbook of Fiber Finish Technology, Marcel Dekker, New York, 1998. [19] Riedo, E., Levy, F., Brune, H., Li, Z.-Q., Kawazoe, Y., and Zhang, S.B., Kinetics of capillary condensation in nanoscopic sliding friction, Phys. Rev. Lett., 88, 185505, 2002. [20] Hansen, W.W. and Tabor, D., Hydrodynamic factors in the friction of fibers and yarns, Textile Res. J., 27, 300–306, 1959. [21] Chapman, J.A., Pascoe, M.W., and Tabor, D., The Friction and Wear of Fibers, Conference on Fiber Friction, Ghent, Belgium, September 1954, pp. P3–P19. [22] Lyne, D.G., The dynamic friction between cellulose acetate yarn and a cylindrical metal surface, Textile Res. J., 46, 112, 1955. [23] Rubenstein, C., The friction and lubrication of yarns, J. Textile Inst., 49, T42, 1958. [24] Rubenstein, C., General theory of the friction of solids, Proc. Phys. Soc., B69, 921, 1956. [25] Rothschild R-2088 Manual and Operating Instructions. [26] ASTM Method D 3412-89 (1991), Standard test method for coefficient of friction, yarn to yarn, Annual Book of ASTM Methods, Volume 07.02, ASTM, Philadelphia, pp. 11–15. [27] ASTM Method D 3103 (1991), Standard test method for coefficient of friction, yarn to solid surfaces, Annual Book of ASTM Methods, Volume 07.01, ASTM, Philadelphia, pp. 848–854. [28] Lawson-Hemphill CTT Operations Manual. [29] Schick, M.J., Friction and lubrication of synthetic fibers, in Surface Characteristics of Fibers and Textiles, Schick, M.J. (Ed.), Marcel Dekker Inc, New York, 1975, p. 11. [30] Prevorsek, D.C. and Sharma, R.K., Fiber–fiber coefficient of friction: effects of modulus and tan δ, J. Appl. Polym. Sci., 23, 173, 1979.

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[31] Pastore, C.M. and Kiekens, P. (Eds), Surface Characteristics of Fibers and Textiles, Volume 94, Surfactant Science Series, Marcel Dekker, New York, 2000. [32] Schlatter, C. and Demas, H.J., Friction studies on Caprolan® yarn, Textile Res. J., 32, 87, 1962. [33] Steinbuch, R. Th., How the crystallization of nylon affects processing and properties, Mod. Plastics, 42, 137, 1964. [34] Berg, J.C., Wettability, Volume 49, Surfactant Science Series, Marcel Dekker, New York, 1993. [35] Myers, D., Surfaces, Interfaces and Colloids, VCH, New York, 1991. [36] Rosen, M.J., Surfactants and Interfacial Phenomena, Second Edition, John Wiley and Sons, New York, 1989. [37] Birdi, K.S., Handbook of Surface and Colloid Chemistry, CRC Press, New York, 1997. [38] Seemuth, P.D. and Potter, J.F., Lubricated Fluoropolymer Yarn, U.S. Patent 6,764,762, July 2004. [39] Doufas, A.K., McHugh, A.J., Miller C. et al., Simulation of melt spinning including flow-induced crystallization — Part II. Quantitative comparisons with industrial spin line data, J. Non-Newton. Fluid, 92, 81–103, 2000. [40] Doufas, A.K. and McHugh, A.J., Simulation of melt spinning including flow-induced crystallization. Part III. Quantitative comparisons with PET spin line data, J. Rheol., 45, 403–420, 2001. [41] Ziabicki, A., Fundamentals of Fiber Formation, Wiley, New York, 1976. [42] Coleman, M.M. and Painter, P.C., Fundamentals of Polymer Science, Second Edition, CRC Press, 1998. [43] Schmack, G., Tandler, B., Optiz, G. et al., High-speed melt spinning of various grades of polylactides, J. Appl. Polym. Sci., 91, 800–806, 2004. [44] Seemuth, P.D. and Anderson, M.B., Bioabsorbable Filaments and Their Production, U.S. Patent 5,288,516, August 1994. [45] Agrawal, A.K. and Bhalla, R., Advances in the production of poly (lactic acid) fibers. A review, J. Macromol. Sci-Polym., R C43, 479–503, 2003. [46] Registered trademark of Bayer. [47] Registered trademark of INVISTA Inc. [48] Registered trademark of E.I. du Pont de Nemours and Co., Inc. [49] Registered trademark of KoSa Industries Inc. [50] Registered trademark of Nylstar®, a joint venture of Rhodia and Snia. [51] Rusznak, I., Handbook of Fiber Science and Technology: Chemical Processing of Fibers and Fabrics, Volume I, Part A, Lewin, A.S. and Sello, S.B. (Eds), Marcel Dekker, New York, 1984. [52] See www.santoni.com for examples of seamless circular knitting machines. [53] Adanur, S. and Mohamed, M.H., Weft insertion on air-jet looms: velocity measurements and influence of yarn structure. Part I: experimental system and computer interface, J. Text. Inst., 2, 297, 1988. [54] Adanur, S. and Mohamed, M.H., Weft insertion on air-jet looms: velocity measurements and influence of yarn structure. Part II: effects of system parameters and yarn structure, J. Text. Inst., 2, 316, 1988. [55] Adanur, S. and Mohamed, M.H., Analysis of yarn motion in single nozzle air-jet filling insertion Part I: theoretical models for yarn motion, J. Text. Inst., 83, 45, 1992. [56] Adanur, S. and Mohamed, M.H., Analysis of yarn motion in single nozzle air-jet filling insertion Part II: experimental validation of the theoretical models and statistical analysis, J. Text. Inst., 83, 56, 1992. [57] Salama, M., Adanur, S., and Mohamed, M.H., Mechanics of a single nozzle air-jet filling insertion system. Part III: yarn insertion through tubes, Textile Res. J., 57, 44, 1987. [58] Adanur, S. and Mohamed, M.H., Analysis of yarn tension in air-jet filling insertion, Textile Res. J., 61, 259, 1991. [59] Adanur, S. and Mohamed, M.H., Analysis of air flow in air-jet filling insertion, Textile Res. J., 61, 253, 1991.

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[60] Bakhtiyarov, S. and Adanur, S., Analysis of air flow in single nozzle air-jet filling insertion: corrugated channel model, Textile Res. J., 66, 401, 1996. [61] Bakhtiyarov, S. and Adanur, S., Airflow over wavy yarn in air-jet filling insertion, Math. Comput. Appl., 4, 1, 1999. [62] Blau, P.J. and Blau, Peter J., Friction Science and Technology, Dekker/CRC Press, New York, 1995. [63] Ikada, Y. and Uyama, Y., Lubricating Polymer Surfaces, CRC Press, 1998. [64] Lansdown, Ar.R., Lubrication and Lubrication Selection, Third Edition, American Society of Mechanical Engineers, 2003. [65] Mang, T. and Dresel, W. (Eds), Lubricants and Lubrication, Wiley-VCH, New York, 2001. [66] Klamann, D., Lubricants and Related Products: Synthesis, Properties, Applications, International Standards, John Wiley and Sons Inc., New York, 1984. [67] Streitwieser, A. and Heathcock, C.H., Introduction to Organic Chemistry, MacMillan, New York, pp. 270, 400, 510, 863, 1985. [68] Hoffman, G.M., Newton, P.E., Birnbaum, H.A., and Kennedy, Jr., G.L., Acute aerosol inhalation studies in several animal species of ethylene oxide/propylene oxide copolymer (Ucon 50-HB-5100), Drug Chem. Toxicol., 14, 243, 1991. [69] Ulrich, C.R., Geil, R.G., Tyler, T.R., and Kennedy Jr., G.L., Two-week aerosol inhalation study in rats of ethylene oxide/propylene oxide copolymers, Drug Chem. Toxicol., 15, 15, 1992. [70] Sjöblom, J. (Ed.), Emulsions and Emulsion Stability, Volume 61, Surfactant Science Series, Marcel Dekker, New York, 1996. [71] Rosano, H.L. and Clausse, M. (Eds), Microemulsion Systems, Volume 24, Surfactant Science Series, Marcel Dekker, New York, 1987. [72] Solans, C. and Kunieda, H. (Eds), Industrial Applications of Microemulsions, Volume 66, Surfactant Science Series, Marcel Dekker, New York, 1996. [73] Datyner, A. (Ed.), Surfactants in Textile Processing, Volume 14, Surfactant Science Series, Marcel Dekker, New York, 1983. [74] Schick, M.J. (Ed.), Nonionic Surfactants, Volume 1, Surfactant Science Series, Marcel Dekker, New York, 1966. [75] Schick, M.J. (Ed.), Nonionic Surfactants: Physical Chemistry, Volume 23, Surfactant Science Series, Marcel Dekker, New York, 1966. [76] van Os, N.M. (Ed.), Nonionic Surfactants: Organic Chemistry, Volume 72, Surfactant Science Series, Marcel Dekker, New York, 1997. [77] Lin, I.J., Friend, J.P., and Zimmels, Y., J. Colloid Interface Sci., 45, 378, 1973. [78] Nace, V.M. (Ed.), Nonionic Surfactants: Polyoxyalkylene Block Copolymers, Volume 60, Surfactant Science Series, Marcel Dekker, New York, 1996. [79] Schmolka, I.R., J. Am. Oil Chem. Soc., 59, 322, 1982. [80] Gloxhuber, C. and Künstler, K. (Eds), Anionic Surfactants: Biochemistry, Toxicology, Dermatology, Second Edition, Revised and Expanded, Volume 43, Surfactant Science Series, Marcel Dekker, New York, 1992. [81] Stache, H.W. (Ed.), Anionic Surfactants: Organic Chemistry, Volume 56, Surfactant Science Series, Marcel Dekker, New York, 1995. [82] Cross, J. (Ed.), Anionic Surfactants: Analytical Chemistry, Second Edition, Revised and Expanded, Volume 73, Surfactant Science Series, Marcel Dekker, New York, 1998. [83] Richmond, J.M. (Ed.), Cationic Surfactants: Organic Chemistry, Volume 34, Surfactant Science Series, Marcel Dekker, New York, 1990. [84] Rubingh, D.N. and Holland, P.M. (Eds), Cationic Surfactants: Physical Chemistry, Volume 37, Surfactant Science Series, Marcel Dekker, New York, 1990. [85] Cross, J. and Singer, E.J. (Eds), Cationic Surfactants: Analytical and Biological Evaluation, Volume 53, Surfactant Science Series, Marcel Dekker, New York, 1990. [86] Becher, P. (Ed.), Encyclopedia of Emulsion Technology, Volume 1, 1983, Volume 2, 1985, Volume 3, 1988, and Volume 4, 1996, Marcel Dekker, New York.

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[87] Hearle, J.W., Moisture and electrical properties, in Moisture in Textiles, Hearle, J.W. and Peters, R.H. (Eds), Butterworths, London, 1960. [88] Grindstaff, T., Improved staple processing performance through air change control, Textile Res. J., 55, 266, 1985. [89] Shenai, V.A., Technology of Textile Processing, Volume V, Sevak Publications, Bombay, 1976. [90] Anderson, N., Peak, R., and Moyse, J.A., Spin Finish with Anti-Static Agent, U.S. Patent 4,294,709, 1981. [91] Schonfeldt, N., Surface Active Ethylene Oxide Adducts, Pergamon Press, Oxford, 1969. [92] Seemuth, P.D., Aqueous Emulsion Finishes for Spandex Fibers Containing Polydimethyl Siloxane and Ethoxylated Long-chain Alkanol, U.S. Patent 4,999,120, 1991. [93] Seemuth, P.D., Chemical, Physical and Structure-Property Relationships of Finish Finishes and Components, 198th National ACS Meeting, August 1989. [94] Seemuth, P.D., Fundamental Interactions of Wetting Phenomena and Finish Compositions, Wetting Fundamental Conference, Charlotte, NC, June 1995. [95] Kissa, E. (Ed.), Fluorinated Surfactants: Synthesis, Properties, Applications, Volume 50, Surfactant Science Series, Marcel Dekker, New York, 1993. [96] Hill, R.A. (Ed.), Silicone Surfactants, Volume 86, Surfactant Science Series, Marcel Dekker, New York, 1999. [97] Patterson, H.T. and Proffitt, Jr., T.J., Fatty acids in textiles, in Fatty Acids in Industry, Johnson, R.W. and Fritz, E. (Eds), Marcel Dekker, New York, 1989, chapter 19. [98] Finch, N.L., Lemley, J.D., and Proffitt, Jr., T.J., High Temperature Resistant Textile Fiber Finish Composition, U.S. Patent 3,544,462, 1970.

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25 Food-Grade Lubricants and the Food Processing Industry 25.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.2 The Food Processing Industry . . . . . . . . . . . . . . . . . . . . . . . . 25.3 Current Registration Practices . . . . . . . . . . . . . . . . . . . . . . . .

25-1 25-2 25-2

U.S. Regulations Prior to 1998 • Changes in Food-Grade Lubrication Standards After 1998 • Third-Party Certifications

25.4 Challenges Facing Food-Grade Lubricants . . . . . . . . . . 25.5 Food-Grade Lubricants Defined by Category . . . . . . . 25.6 Approved Lubricant Formulations in H1 Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

25-4 25-4 25-5

Acceptable Food-Grade Basestocks • Acceptable Food-Grade Additives and Thickeners

James C. Fitch, Sabrin Gebarin, and Martin Williamson Noria Corporation

25.7 Selecting What Machines Require Food-Grade Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.8 Selecting an H1 Food-Grade Supplier [13] . . . . . . . . . . 25.9 Global Trends [2] . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.10 Religious Organizations Influence in Food-Grade Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.11 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

25-12 25-12 25-14 25-17 25-17 25-17

25.1 Introduction The food processing industry presents unique challenges to lubricant formulation engineers, lubricant marketers, plant lubrication engineers, equipment designers, and builders. While it is never desirable for lubricants to be allowed to contaminate raw materials, work-in-progress, or finished product, the consequences of a lubricant contaminated product is rarely more acute than in the food processing industry. As such, lubricants used in this industry have requirements, protocols, and performance expectations that go well beyond typical industrial lubricants.

25-1

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This chapter provides a general overview of the unique challenges associated with food-grade lubricants including recent revisions of the regulatory environment. The terminology commonly used by suppliers and clients of food-grade products will also be defined and discussed. So too, will be machine applications common to certain sectors of the food processing industry and their unique requirements for food-grade lubricants.

25.2 The Food Processing Industry Food-grade lubricants are significant in scope and application when you consider the size of the food processing industry. In the United States, food manufacturers represent a significant percentage of total manufacturing. According to the 1997 U.S. Census, approximately 485 billion dollars in sales revenue was generated (about the same amount generated in the transportation manufacturing industry). This represents almost 13% of all manufacturing in the United States. In 1997, approximately 28,000 manufacturing facilities employed 1.6 billion employees and produced 233 billion dollars in goods [1].

25.3 Current Registration Practices Historically, the two U.S. government agencies primarily involved in food processing were the United States Department of Agriculture (USDA) which regulates meat, poultry, and plants and the United States Food and Drug Administration (FDA) which monitors other food and pharmaceutical manufacturing operations.

25.3.1 U.S. Regulations Prior to 1998 Prior to 1998, approval and compliance of food-grade lubricants was the responsibility of the USDA. The Food Safety and Inspection Services (FSIS), headed by the USDA, reviewed the formulations of maintenance and operating chemicals. FSIS required meat and poultry facilities to use only nonfood compounds that were pre-approved by the USDA authorization program. However, these programs spread to other food market sectors such as fisheries and retail food operations [2]. To gain USDA approval, lubricant manufacturers had to prove that all the ingredients in the formulation were allowable substances. Allowable substances, in this instance, are those listed by the FDA in accordance with the Guidelines of Security Code of Federal Regulations (CFR) Title 21, §178.3570. This did not include lubricant testing. Rather, the approval was based primarily on a review of the formulation ingredients of the lubricant [2].

25.3.2 Changes in Food-Grade Lubrication Standards After 1998 Starting February 1998, FSIS significantly altered their program by implementing a system established by Hazard Analysis and Critical Control Point (HACCP) requiring the manufacturer to assess risk at each point in the operation where contamination might occur. The National Aeronautics and Space Administration (NASA) originally developed the HACCP system in the 1960s to prevent astronauts from receiving any food-borne illnesses. It established measures like minimum cooking temperatures for each control point, instituted procedures to monitor these measures, and also provides corrective actions if critical limits are not met [3]. In essence, the manufacturer became responsible for reviewing and approving the chemical compositions of lubricants to decide whether they were safe or not as food-grade lubricants.

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Food-Grade Lubricants and the Food Processing Industry

25-3

FIGURE 25.1 NSF’s search engine of approved lubricants in H1, H2, and H3 applications.

25.3.3 Third-Party Certifications In response to the change in the approval process, several commercial organizations developed external certification programs. Three such organizations were the National Sanitation Foundation (NSF), Underwriters Laboratory (UL), and a joint effort by three recognized industry professional associations: The National Lubricating Grease Institute (NLGI), The European Lubricating Grease Institute (ELGI), and the European Hygienic Equipment Design Group (EHEDG). NSF has developed a lubricant evaluation program that essentially mirrors the FSIS program by evaluating the candidate lubricant formulations to verify compliance with the various FDA CFR guidelines. Each component in the formulation is submitted to NSF by the lubricant manufacturer along with other supporting documentation. This is then reviewed to verify it is within the FDA list of permitted substances [4]. NSF’s website provides food processing manufacturers with a continually updated list of approved lubricants at www.nsfwhitebook.org (Figure 25.1). Underwriters Laboratory is another organization that began third-party certification of food-grade lubricants but no longer is doing so. While they have not been as active as NSF in the area of food-grade lubricants, in the past, UL has organized several informational meetings inviting lubricant and chemical manufacturers to attend [5]. The NLGI/ELGI/EHEDG Joint Food-grade Lubricants Working Group has been active in drafting an authorization program for food-grade lubricants. This group’s program is also based on the former USDA/FSIS authorization program and CFR policies. Their plan is to develop a DIN (the German Institute for Standardization) standard in Germany and use the DIN standard to later develop an ISO (International Organization for Standardization) standard [5]. Not all countries use third-party certifications. Canada, New Zealand, Australia, and Japan are some of the countries that federally regulate food-grade lubricants [1]. However, the Canadian Food Inspection Agency (CFIA) is working on a food-grade lubricants approval system, and NSF will help with the

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CFIA review process. Also, the Australian Quarantine Inspection Service has approved approximately 50 food-grade lubricants based on NSF registration [6].

25.4 Challenges Facing Food-Grade Lubricants Agricultural and animal substances go through a number of processes in a manufacturing plant such as cleansing, sterilizing, blending, mixing, cooking, freezing, cutting, packaging, canning, and bottling. Large-scale food processing requires machinery such as pumps, mixers, tanks, hoses, and pipes, chain drives, and conveyor belts. Machinery used in food processing facilities face many of the same tribological and lubrication challenges found in other nonfood processing plants. In that sense, lubricants must offer similar protection of internal surfaces to control friction, wear, corrosion, heat, and deposits. They must also offer good pumpability, oxidation stability, hydrolytic stability, and thermal stability where the application so requires. Many of the raw materials used to formulate lubricants that effectively address these challenges in conventional industrial applications are not permissible in food applications for safety reasons. In addition, certain applications within the food and drug manufacturing facilities demand that lubricants resist degradation and impaired performance when in contact with food products, certain process chemicals, water (including steam), and bacteria. They must also exhibit neutral behavior toward plastics and elastomers and have the ability to dissolve sugars. In general, these lubricants must comply with food/health and safety regulations, as well as be physiologically inert, tasteless, odorless, and internationally approved [7]. Lubricants in many food processing plants can be subjected to ingression and contend with an assortment of environmental contaminants. For instance, a corn-milling environment generates significant dust. Although not as hard as silica-based terrain dust, it still presents a problem for filtration. A meat plant requires stringent steam cleaning at all times, so the risk of water contamination is high. Water contamination in gear oils routinely exceeds 15% in some plants. Another aspect of lubrication contamination that poses unique risk to food-grade lubricants is the growth of microorganisms such as bacteria, yeast, and fungi. While these can also be challenging to conventional industrial lubricants, the opportunity for microbial contamination in the food-production industry is considerably greater.

25.5 Food-Grade Lubricants Defined by Category Food-grade lubricants are either compounded or uncompounded products acceptable for use in meat, poultry, and other food processing equipment, applications, and plants. The lubricant types in food-grade applications are broken into categories based on the likelihood that they will contact food. The original food-grade designations H1, H2, and H3 were created by the USDA. The approval and registration of a new lubricant into one of these categories depends on the ingredients used in the formulation. The three designations are described here [2]. H1 lubricants are food-grade lubricants used in food-processing environments where there is some possibility of incidental food contact. Lubricant formulations must be composed of one or more approved basestocks, additives, and thickeners (if grease) listed in 21 CFR 178.3750. Only the minimum amount of lubricant required should be used on the equipment. H2 lubricants are lubricants used on equipment and machine parts in locations where there is no possibility that the lubricant or lubricated surface contacts food. Because there is not risk of contacting food, H2 lubricants do not have a defined list of acceptable ingredients. They cannot, however, contain intentionally heavy metals such as antimony, arsenic, cadmium, lead, mercury or selenium. Also, the ingredients must not include substances that are carcinogens, mutagens, teratogens, or mineral acids [4].

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Food-Grade Lubricants and the Food Processing Industry TABLE 25.1

25-5

H-3 Soluble Oil Approved Lubricants

Lubricant type

Regulations they must meet

Edible oils (corn oils, cottonseed oil, soybean oil) Certain mineral oils Generally recognized as safe (GRAS)

21 CFR 172.860 21 CFR 172.878 21 CFR 182 or 21 CFR 184

H3 lubricants, also known as soluble or edible oil, are used to clean and prevent rust on hooks, trolleys, and similar equipment. Equipment applied with H3 lubricants should be cleaned by washing or wiping the surface before putting the equipment in service. These lubricants can only consist of ingredients as shown in Table 25.1 [4]. Deciding whether there is a possibility of contact is tough, and many have erred on the side of safety with respect to selecting H1 over H2.

25.6 Approved Lubricant Formulations in H1 Lubricants As previously mentioned, the USDA/FSIS approvals are based on the various FDA Codes in Title 21 that dictate approval for ingredients used in lubricants that may have incidental contact with food. These are mentioned as follows: • 21.CFR 178.3570 — allowed ingredients for the manufacture of H1 lubricants • 21.CFR 178.3620 — white mineral oil as a component of nonfood articles intended for use in contact with food • 21.CFR 172.878 — USP mineral oil for direct contact with food • 21 CFR 172.882 — synthetic isoparaffinic hydrocarbons • 21.CFR 182 — substances generally recognized as safe Based on the Title 21 FDA regulations noted, the following is paragraphs discuss the allowable basestocks, additives, and thickeners in food-grade lubricants.

25.6.1 Acceptable Food-Grade Basestocks Depending on whether the food-grade lubricant is H1 or H2, the list of approved basestocks will vary. H2 lubricant basestock guidelines are less restrictive and consequently allow a broader variety of basestocks. Many products used in industrial (nonfood) plants are also used in food plants for H2 applications. H1 lubricants are much more limited since they are designed to allow for accidental exposure with the processed foods. The approved H1 lubricant basestocks can be either mineral or synthetic. 25.6.1.1 Petroleum-Based Lubricants Mineral oils used in H1 food-grade lubricants are either technical white mineral or USP-type white mineral oils. White oils start as normal paraffinic petroleum stocks and are processed into pure branched paraffin stocks, stripped free of the majority of aromatic hydrocarbons, sulfur, and nitrogen contaminants. They are highly refined and are colorless, tasteless, odorless, and nonstaining. Technical white oils meet the regulations specified in 21 CFR 178.3620. Based on the American Society for Testing Materials (ASTM) method D156-82, “Standard Test Method for Saybolt Color of Petroleum Products (Saybolt Chromometer Method),” the Saybolt color must be a minimum of 20 to be considered a technical white oil [8]. USP mineral oils are the purest of all white mineral oils, and are the most oxidatively stable [5]. Historically, white mineral oils were first listed in the United States Pharmacopoeia (USP) in 1926. Later, a paper on the general principles of white oil manufacturing was written in 1935 followed by other papers [1].

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TABLE 25.2 Approved Substance for H1 Lubricants per 21CFR 178.3570 Substance

Limitations

Aluminum stearoyl benzoyl hydroxide

For use only as a thickening agent in mineral oil lubricants at a level not to exceed 10% by weight of the mineral oil

N , N -bis(2-ethylhexyl)-armethyl-1H-benzotriazole-1methanamine (CAS Reg. No.94270-86-7) BHA BHT

For use as a copper deactivator at a level not to exceed 0.1% by weight of the lubricant

[alpha]-Butyl-omega- hydroxypoly(oxyethylene) poly(oxypropylene) produced by random condensation of a 1 : 1 mixture by weight of ethylene oxide and propylene oxide with butanol; minimum mol wt 1500; Chemical Abstracts Service Registry No. 9038-95-3 [alpha]-Butyl-omega- hydroxypoly (oxypropylene); minimum mol wt 1,500; Chemical Abstracts Service Registry No. 9003-13-8 Castor oil

Addition to food not to exceed 10 ppm

Castor oil, dehydrated

Addition to food not to exceed 10 ppm

Castor oil, partially dehydrated

Addition to food not to exceed 10 ppm

Dialkyldimethylammonium aluminum silicate (CAS Reg. No. 68953-58-2), weight 1, 6hexanediol (CAS Reg. No. by weight of the mineral oil. 629-11-8), where the alkyl groups are derived from hydrogenated tallow fatty acids (C14 –C18 ) and where the aluminum silicate is derived from bentonite Dimethylpolysiloxane (viscosity greater than 300 cSt)

For use only as a gelling agent in mineral oil lubricants at a which may contain up to 7% by level not to exceed 15%

Addition to food not to exceed 10 ppm

Addition to food not to exceed 10 ppm

Addition to food not to exceed 1 ppm

Di (n-octyl) phosphite (CAS Reg. No. 1809-14-9)

For use only as an extreme pressure-antiwear adjuvant at a level not to exceed 0.5% by weight of the lubricant

Disodium decanedioate (CAS Reg. No. 1726514-4)

For use only:

Disodium EDTA (CAS Reg. No. 139-33-3)

For use only as a chelating agent and sequestrant at a level not to exceed 0.06% by weight of lubricant at final use dilution

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1. As a corrosion inhibitor or rust preventative in mineral oil–bentonite lubricants at a level not to exceed 2% by weight of the grease 2. As a corrosion inhibitor or rust preventative only in greases at a level not to exceed 2% by weight of the grease

Food-Grade Lubricants and the Food Processing Industry

TABLE 25.2

25-7

Continued

Substance Ethoxylated resin phosphate ester mixture consisting of the following compounds: 1. Poly(methylene-p-tert-butyl phenoxy)poly(oxyethylene) mixture of dihydrogen phosphate and monohydrogen phosphate esters (0 to 40% of the mixture). The resin is formed by condensation of 1 mol of p-tertbutylphenol with 2 to 4 mols of formaldehyde and subsequent ethoxylation with 4 to 12 mols of ethylene oxide 2. Poly(methylene-p-nonylphenoxy) poly(oxyethylene) mixture of dihydrogen phosphate and monohydrogen phosphate esters (0 to 40% of the mixture). The resin is formed by condensation of 1 mol of p-nonylphenol with 2 to 4 mols of formaldehyde and subsequent ethoxylation with 4 to 12 mols of ethylene oxide 3. n-Tridecyl alcohol mixture of dihydrogen phosphate and monohydrogen phosphate esters (40 to 80% of the mixture; CAS Reg. No. 56831-62-0)

Limitations For use only as a surfactant to improve lubricity in lubricating fluids complying with this section at a level not to exceed 5% by weight of the lubricating fluid

Fatty acids derived from animal or vegetable sources and the hydrogenated forms of such fatty acids 2-(8-Heptadecenyl)-4,5-dihydro-1H-imidazole-1ethanol (CAS Reg. No. 95-38-5)

For use at levels not to exceed 0.5% by weight of the lubricant

Hexamethylenebis(3,5-di-tert-butyl-4-hydroxy hydrocinnamate) (CAS Reg. No.35074-77-2)

For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant

[alpha]-Hydro-omega-hydroxypoly (oxyethylene) poly(oxypropylene) produced by random condensation of mixtures of ethylene oxide and propylene oxide containing 25 to 75% by weight of ethylene oxide; minimum mol wt 1,500; Chemical Abstracts Service Registry No. 9003-11-6 12-Hydroxystearic acid

Addition to food not to exceed 10 ppm

Isopropyloleate

For use only as an adjuvant (to improve lubricity) in mineral oil lubricants

Magnesium ricinoleate

For use only as an adjuvant in mineral oil lubricants at a level not to exceed 10% by weight of the mineral oil

Mineral oil

Addition to food not to exceed 10 ppm

N-Methyl-N-(1-oxo-9-octadecenyl) glycine (CAS Reg. No. 110-25-8)

For use as a corrosion inhibitor at levels not to exceed 0.5% by weight of the lubricant

N-phenylbenzenamine, reaction products with 2, 4, 4-trimethylpentene (CAS Reg. No. 68411-46-1)

For use only as an antioxidant at levels not to exceed 0.5% by weight of the lubricant

Petrolatum

Complying with Sec. 178.3700. Addition to food not to exceed 10 ppm

Phenyl-[alpha]-and/or phenyl [beta]-naphthylamine

For use only, singly or in combination, as antioxidant in mineral oil lubricants at a level not to exceed a total of 1% by weight of the mineral oil (Continued)

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TABLE 25.2

Continued

Substance

Limitations

Phosphoric acid, monohexyl and dihexyl esters, compounds with tetramethylnonylamines and C11−14 alkylamines Phosphoric acid, monoisooctyl and diisooctyl esters, reacted with tert-alkyl and (C12 –C14 ) primary amines (CAS Reg. No.68187-67-7)

For use only as an adjuvant at levels not to exceed 0.5% by weight of the lubricant

Phosphorothioic acid, O, O, O-triphenyl ester, tertbutyl derivatives (CAS Reg. No. 192268-65-8)

For use only as an extreme pressure-antiwear adjuvant at a level not to exceed 0.5% by weight of the lubricant

Polyurea, having a nitrogen content of 9 to 14% based on the dry polyurea weight, produced by reacting tolylene diisocyanate with tall oil fatty acid(C16 and C18 ) amine and ethylene diamine in a 2 : 2 : 1 molar ratio Polybutene (minimum average mol wt 80,000)

For use only as an adjuvant in percent level not to exceed 10 mineral oil lubricants at a by weight of the mineral oil

For use only as a corrosion inhibitor or rust preventative in lubricants at a level not to exceed 0.5% by weight of the lubricant

Addition to food not to exceed 10 ppm

Polybutene, hydrogenated; complying with the identity prescribed under Sec. 178.3740 Polyethylene

Addition to food not to exceed 10 ppm

Polyisobutylene (average mol wt 35,000–140,000 (Flory))

For use only as a thickening agent in mineral oil lubricants

Sodium nitrite

Use only as a rust preventive in mineral oil lubricants at a level not to exceed 3% by weight of the mineral oil

Tetrakis[methylene(3,5-di-tert-butyl-4hydroxyhydro-cinnamate)] methane (CAS Reg. No. 6683-19-8) Thiodiethylenebis (3,5-di-tert-butyl-4- hydroxyhydrocinnamate) (CAS Reg. No. 41484-35-9)

For use only as an antioxidant in lubricants at a level not to exceed 0.5% by weight of the lubricant

Tri[2(or 4)-C9−−10 -branched alkylphenyl] phosphorothioate (CAS Reg. No. 126019-82-7)

Triphenyl phosphorothionate (CAS Reg. No. 597-82-0) Tris(2,4-di-tert-butylphenyl)phosphite (CAS Reg. NO. 31570-04-4) Thiodiethylenebis(3,5-di-tert-butyl-4-hydroxyhydro-cinnamate) (CAS Reg. No. 41484-35-9) Zinc sulfide

Addition to food not to exceed 10 ppm

For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant For use only as an extreme pressure-antiwear adjuvant at levels not to exceed 0.5% by weight of the lubricant For use as an adjuvant in lubricants herein listed at a level not to exceed 0.5% by weight of the lubricant For use only as a stabilizer at levels not to exceed 0.5% by weight of the lubricant For use as an antioxidant at levels not to exceed 0.5% by weight of the lubricant For use at levels not to exceed 10% by weight of the lubricant

Source: 21 CFR 3570 — Lubricants with incidental food contact. Retrieved online at www.access.gpo.gov/ nara/cfr.index.asp.

25.6.1.2 Synthetic Lubricants Synthetic H1 lubricants are mainly polyalphaolefins (PAO). They were first introduced in 1981 by Gulf Research and Development Company [1]. Compared to white mineral oils, they have significantly greater oxidation stability and greater range of operating temperatures. Another H1 synthetic lubricant used is Polyalkylene glycols (PAG). These lubricants are more increasingly used in high temperature applications.

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Food-Grade Lubricants and the Food Processing Industry TABLE 25.3

25-9

Bakery and Confectionery Products

Flour pumps/blowers/fluidizers Greased bearings NLGI No. 2 Grease Oiled bearings ISO R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Blenders Greased bearings Enclosed gears

Blanchers — initial cooking Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Shapers/rounders/moulders/elongators Geared bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray

Dough mixers Greased bearings Oiled bearings Enclosed gears Chains Hydraulics

NLGI No. 2 Grease ISO R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Panners/depanners Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray

Sifters/separators Greased bearings Chains Air Line Lube

NLGI No. 2 Grease NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Proofers/coolers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray

Dividers/portioners Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray

Pan-tray washers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil

Ovens/fryer/roasters/cookers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray Baggers/packagers Greased bearings Oiled bearings Open gears Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray

Pan-tray stackers/unstackers Greased bearings NLGI No. 2 Grease Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil

Wrappers Greased bearings Oiled bearings Open gears Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray

Liquor mills Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 220 Gear Oil

Dust collectors Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 100 R&O or AW Oil

Depositors/applicators Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Conveyers Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray

Extruders Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Pumps Greased bearings

NLGI No. 2 Grease

Slicers Greased bearings Enclosed gears Chains Air Line Lube

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray or ISO 100 R&O or AW Oil

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

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TABLE 25.4

Beverages

Depalletizers/palletizers Greased bearings Enclosed gears Chains Hydraulics Air Line Lubes

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil ISO 46 R&O or AW Oil

Bottle-can uncaser/packer/case wrapper Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil

Bottle-can washers/rinsers Greased bearings Enclosed gears Chains Open gears

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray

Bottle-can fillers Greased bearings Oiled bearings Enclosed gears

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil

Pasteurizers Greased bearings Enclosed gears Hydraulics Chains

NLGI No. 2 Grease ISO 460 Gear Oil ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray

Bottle cappers Greased bearings Enclosed gears Thread roller

NLGI No. 2 Grease ISO 460 Gear Oil ISO 100 R&O or AW Oil

Conveyors Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray

Can closers Greased bearings Enclosed gears Chains Open gears

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray

Pumps Greased bearings

NLGI No. 2 Grease

Bottle labelers Greased bearings/cams Oiled bearings Chains Open gears Enclosed gears

NLGI No 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray NLGI No. 2 Grease Spray ISO 220 Gear Oil

Syrup mixers/pumps/fillers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Open gears NLGI No. 2 Grease Spray

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

Dimethylpolysiloxane (silicones) with a viscosity greater than 300 cSt [9] is also permitted for H1 lubricants. Sanction letters for the use of silicone fluids as defoaming agents show up as early as 1953. Silicones were not approved until soon after a petition filed by General Electric in 1965 [1]. Silicones have even higher thermal and oxidation stability than PAO and PAG base oils. 25.6.1.3 Differences Among Basestocks Although synthetics are more expensive than mineral oils, tests performed on H1 PAO and white mineral oils on drive chains show that the useful life of PAOs is almost twice that of white oils. Testing has shown PAG base oils have a service life five times longer than white mineral oils [7]. In addition to longer service life, there is evidence that synthetic H1 oils do a better job of protecting metal surfaces from corrosion and wear and withstand greater temperature extremes required around freezers or ovens.

25.6.2 Acceptable Food-Grade Additives and Thickeners Often basestocks are not able to meet the severe demands required in food processing work environments. To improve the performance characteristics of base oils, additives are blended into the formulation. The types of antioxidants, corrosion inhibitors, antiwear, extreme pressure additives, and concentration are limited by 21 CFR 178.3570.

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Food-Grade Lubricants and the Food Processing Industry TABLE 25.5

25-11

Canned, Preserved, and Frozen Fruits and Vegetables

Peelers/pitters/huskers Greased bearings/slides Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray

Centrifuges Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray

Snippers Greased bearings/slides Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray

Presses Geared bearings Enclosed gears Chains Hydraulics

NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Graders/food washers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray

Cookers/coolers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray

Blanchers — initial cooking Greased bearings NLGI No. 2 Grease Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray

Freezing tunnels Greased bearings Hydraulics Chains

NLGI No. 2 Grease ISO 46 R&O or AW Oil NLGI No. 2 Grease Spray

Grinders/blenders/finishers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Vacuum filters/evaporators Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray

Hammer mill Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 100 R&O or AW Oil

Homogenizers/pasteurizers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil or ISO 460 Gear Oil Fluid drive ISO 46 R&O or AW Oil Hydraulics ISO 68 R&O or AW Oil

Baggers/sealers Greased bearings Oiled bearings Enclosed gears Chains/open gears Air Line Lubes

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 100 R&O or AW Oil

Depalletizers/palletizers Greased bearings Enclosed gears Chains Hydraulics Air Line Lubes

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil ISO 46 R&O or AW Oil

Can-jar washers Greased bearings Enclosed gears Chains/open gears Can-jar filters Greased bearings/valves Oiled bearings/valves Enclosed gears

Can closers Greased bearings Enclosed gears Chains/open gears

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray

Labelers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray

Can-jar packers/casers/uncasers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil

Conveyers Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 460 Gear Oil NLGI No. 2 Grease Spray (Continued)

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Continued

Can-jar coolers Greased bearings Enclosed gears Chains/open gears

NLGI No. 2 Grease ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray

Jar cappers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Pumps Greased bearings

NLGI No. 2 Grease

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

Greases are lubricating oils that have a thickening agent added to the formulation. Approved grease thickeners include aluminum stearate, aluminum complex, organo clay, and polyurea [10]. Aluminum complex is the most common H1 food-grade grease thickener. They can withstand high temperatures and are water resistant, which are important properties for food processing applications. Greases with calcium sulfonate thickeners have not been explored for approval by the USDA or FDA, but has been approved in Canada for incidental contact [11]. The list of approved base oils, additives, and thickeners for H1 incidental contact with food is available in Table 25.2.

25.7 Selecting What Machines Require Food-Grade Lubricants Selecting whether to use an H1 or H2 lubricant can be challenging. As previously mentioned, H1 lubricants are permitted where incidental contact might be possible, whereas an H2 lubricant is only permitted where there is no possible contact with the food product. For example, a lubricant used on a conveyor system running over a food line must be an H1 category oil, while a conveyor system running underneath a food line may use either an H1 or H2 lubricant. Because H1 lubricants are limited by types of additives and in the past only used mineral oil basestocks, H1 lubricants in certain instances provided less protection and shorter lubricant life. Now that synthetics are used, some H1 lubricant performance can exceed nonfood-grade lubricants. This is highly significant in allowing consolidation and avoiding accidental cross-contamination of H1 and H2 oils, and contamination of H2 oils with food [7]. Tables 25.3 through 25.9 are designed as a quick reference for some food processing applications generic for several types of industries [12]. The specific application should be checked to verify the lubricant grade or viscosity. The tables do not identify whether to use an H1 or H2 lubricant. It is ultimately the food-processing plant’s decision to determine whether an H1 is required or if an H2 lubricant is allowable.

25.8 Selecting an H1 Food-Grade Supplier [13] Finding the right lubricant supplier is as important as selecting the right lubricant. It is important to find a food-grade lubricant supplier that understands specific applications and requirements. Also, a supplier can serve as a part of the maintenance department, to help educate staff on lubrication maintenance and provide training to get the most performance and service life possible out of the lubricant. Other important qualities of a lubricant supplier are product consolidation, oil analysis, on-time delivery, speedy response to questions, and ability to tailor products to client needs.

© 2006 by Taylor & Francis Group, LLC

Food-Grade Lubricants and the Food Processing Industry TABLE 25.6

25-13

Dairy Products

Separators/clarifiers Enclosed gears

ISO 100 R&O or AW Oil or ISO 220 Gear Oil

Labelers Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 220 Gear Oil

Homogenizers/pasteurizers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil or ISO 220 Gear Oil Hydraulics ISO 46 R&O or AW Oil

Casers/packers/stackers/destackers Geared bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Guides NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil

Tank-vat agitators Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 220 Gear Oil

Cheese fillers/presses Greased bearings Oiled bearings Slides Cams

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil NLGI No. 2 Grease Spray

Fillers/cappers Geared bearings Enclosed gears Chains/open gears Air Line Lube

NLGI No. 0 or 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Butter churns/boats Greased bearings Enclosed gears Hydraulics Rear leg

NLGI No. 2 Grease ISO 220 Gear Oil ISO 46 R&O or AW Oil ISO 68 Turbine Oil

Packagers Geared bearings Enclosed gears Chains/open gears Air Line Lube

NLGI No. 0 or 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Centrifuges Oiled bearings

ISO 100 R&O or AW Oil

Dryers Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 220 Gear Oil

Fruit feeders Hydraulics

ISO 46 R&O or AW Oil

Powder baggers/bag closers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil

Ice cream freezer Greased bearings Oiled bearings Enclosed gears Chains Valves

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Mixers/hammers/mills/vibrators Greased bearings NLGI No. 2 Grease Enclosed gears ISO 100 R&O or AW Oil

Ice cream fillers Greased bearings Slides Oiled bearings Clutches

NLGI No. 2 Grease NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 68 Turbine Oil

Dust collectors Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 100 R&O or AW Oil

Conveyers Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Liquifiers Greased bearings

NLGI No. 2 Grease

Pumps Greased bearings

NLGI No. 2 Grease

Thermutators Greased bearings

NLGI No. 2 Grease

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

© 2006 by Taylor & Francis Group, LLC

25-14

Handbook of Lubrication and Tribology

TABLE 25.7

Fat and Oil Products

Expellers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray

Centrifuges Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Bean cleaners/shakers/dehullers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Mixers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Dryers/condensers/coolers/toasters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chain/open gears NLGI No. 2 Grease Spray

Dust collectors Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Crackers/grinders/hammer mills Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Conveyors Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Flakers Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Pumps Greased bearings

NLGI No. 2 Grease

Solvent extractors Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

25.9 Global Trends [2] USDA H1 and H2 still stand as a recognized approval for food and drug suitability. In fact, many lubricant manufacturers still aspire to the USDA H1 and H2 categories and approval process, and supply certification from their boards of directors to guarantee that claim. However, efforts chaired by Klüber Lubricants of Germany led to the creation of a new standard, DIN V 0010517, 2000-08 (Food-grade Lubricants — Definitions and Requirements). This standard has since been approved at a higher DIN level. This German standard has been submitted by DIN as a draft to ISO in Geneva. It may take several years from the date the application is accepted for an international standard to be released. NSF has evolved globally to succeed the USDA. NSF International, The Public Health and Safety Company™ , has been committed to public health, safety, and protection of the environment for more than 55 years. NSF has earned the Collaborating Center designations by the World Health Organization (WHO) for both food safety and for drinking water safety and treatment. It is conceived and administered as a public service organization serving as an independent and neutral body to resolve issues between regulatory bodies, business, industry, and the public. The DIN standard V 0010517, 2000-08 has also been adopted by ELGI and NLGI as their guideline.

© 2006 by Taylor & Francis Group, LLC

Food-Grade Lubricants and the Food Processing Industry TABLE 25.8

25-15

Grain Mill Products

Milling separatos/degerminators Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

Tempering bins Greased bearings Enclosed gears

Shakers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Grinding mills/hammer mills/comminuters/crumblizers Geared bearings NLGI No. 2 Grease Enclosed gears ISO 220 Gear Oil Adjusting screws NLGI No. 2 Grease Spray

Washers/cleaners Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Dryers Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Evaporators Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Ovens/cookers Greased bearings Enclosed gears

NLGI No. 2 Grease ISO 220 or 460 Gear Oil

Chains

NLGI No. 2 Grease Spray

NLGI No. 2 Grease ISO 220 Gear Oil

Sifters Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Loaf molders/extruders Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Dough mixers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Centrifuges/filters/oil extractors Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil or ISO 220 Gear Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray

Package fillers/baggers/bag closers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 46 R&O or AW Oil Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray

Dust collectors Greased bearings

Mixers/blenders Greased bearings Enclosed gears Chains

Conveyors Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Pellet mills — feed processing Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 100 R&O or AW Oil ISO 220 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Coolers Greased bearings Enclosed gears Chains

Pumps Greased bearings

NLGI No. 2 Grease

NLGI No. 2 Grease

NLGI No. 2 Grease ISO 460 Gear Oil NLGI No. 2 Grease Spray

Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With Permission.

© 2006 by Taylor & Francis Group, LLC

25-16

Handbook of Lubrication and Tribology TABLE 25.9

Meat, Seafood, and Poultry

Parts washers/scalders Greased bearings NLGI No. 2 Grease

Mixers/mincers Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil

Feather pickers

Meat saws/meat and bacon slicers/peelers/skinners/ chippers/venters Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 460 Gear Oil Chains NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil

Greased bearings

NLGI No. 2 Grease

Smoke houses/ovens Greased bearings NLGI No. 2 Grease Linkage NLGI No. 2 Grease Spray Open gears NLGI No. 2 Grease Spray

Sausage linkers/frank machines/patty machines Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 100 R&O or AW oil or 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray

Cookers Greased bearings Linkage

NLGI No. 2 Grease NLGI No. 2 Grease Spray

Stuffers Greased bearings Oiled bearings Enclosed gears

Open gears

NLGI No. 2 Grease Spray

Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 100 R&O or AW Oil or 220 or 460 Gear Oil NLGI No. 2 Grease Spray

Grinders/disintegrators Greased bearings NLGI No. 2 Grease Oiled bearings ISO 100 R&O or AW Oil Enclosed gears ISO 220 Gear Oil Chains/open gears NLGI No. 2 Grease Spray Air Line Lube ISO 46 R&O or AW Oil

Centrifuges/separators/dryers/filters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Air Line Lube ISO 46 R&O or AW Oil

Pickling injectors Greased bearings Enclosed gears Chains Hydraulics

Graders/deboners Greased bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 220 or 460 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Can washers, fillers and closers/labelers/packers/ wrappers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Chains NLGI No. 2 Grease Spray Hydraulics ISO 46 R&O or AW Oil Air Line Lube ISO 46 R&O or AW Oil

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray

Freezers Greased bearings Enclosed gears Chains Hydraulics

NLGI No. 2 Grease ISO 220 Gear Oil NLGI No. 2 Grease Spray ISO 46 R&O or AW Oil

Rehangers/neck breakers Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil

Conveyors Greased bearings Oiled bearings Enclosed gears Chains

NLGI No. 2 Grease ISO 100 R&O or AW Oil ISO 220 Gear Oil NLGI No. 2 Grease Spray

Eviscerators/remove entrails Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil

Pumps Greased bearings

NLGI No. 2 Grease

© 2006 by Taylor & Francis Group, LLC

Food-Grade Lubricants and the Food Processing Industry

25-17

TABLE 25.9 Continued Gizzard machines/lung pullers/neck skin cutters/venters Greased bearings NLGI No. 2 Grease Enclosed gears ISO 220 or 460 Gear Oil Hydraulics ISO 46 R&O or AW Oil Source: Food Processing Industry brochure. Lubrication Engineers, Inc. With permission.

25.10 Religious Organizations Influence in Food-Grade Lubricants The Muslim and Jewish religions further restrict the formulation of food-grade lubricants. Today, there are approximately 14 million Jews and 1.3 billion Muslims worldwide [14]. Both religions have rules covering aspects of food processing. “Kosher for Pareve,” or Kosher, is the term used to describe Jewish dietary laws. Kosher law is approved by several rabbinic orders. In the United States, the Orthodox Union and the Organized Kashrus Laboratories are major approval organizations active in the approval of food-grade lubricants. Kosher law outlaws the use of pork and pork by-products. Kosher law also prohibits any mixing of meats and dairy and eggs. Any equipment must be properly cleaned and left idle for 24 h before and after making kosher foods [1]. Under Islamic law, “Halal,” meaning lawful or permitted in Arabic, laws are imposed on their food products. In the United States, the Islamic Food and Nutrition Council of America issues Halal Certificates. Similar to Kosher laws, Halal foods exclude the use of pork and pork by-products. Also, Halal excludes the use of alcohol in its products, which potentially limits some of the additives used in food-grade lubricants [1].

25.11 Conclusions The food and beverage processing industries with respect to food-grade lubricants has changed dramatically within the last five years. Understanding the differences between H1, H2, and H3 lubricants and making the proper lubricant selection is critical to food safety and machine reliability. As an additional source, NSF’s website provides lubricant requirements for food-grade products and gives a free access listing of certified food-grade lubricants on their website at www.nsfwhitebook.org.

References [1] Raab, Michael J., Food-grade lubricants: a new world order, NLGI Spokesman, 66, 2, 2002. [2] Williamson, M., Understanding food-grade lubricants. Machinery Lubrication Magazine, 64, 2003. [3] Hodson, D., Food-grade lubricants reduce contamination threats for food and beverage manufacturers. Machinery Lubrication Magazine, 24, 2004. [4] NSF International Registration Guidelines (July 2003) version 3.3, Retrieved October 2004 from http://www.nsf.org/business/nonfood_compounds/guidelines.pdf. [5] Girard, J., The continuing evolution of food-grade lubricants. Machinery Lubrication Magazine, 20, 2002. [6] Email correspondence with Dr. Kenji Yano, program manager for NSF International Nonfood Compounds Registration Program. [7] Lauer, D.A., Special lubricants for the food-processing and pharmaceutical industries. Lubrication Excellence 2003 Conference Proceedings, 439, 2003. [8] 21 CFR 178.3620 — Technical White Mineral Oil as a Component of Nonfood Articles Intended For The Use of Contact with Food. Retrieved November 2004 at http://www.gpoaccess.gov/cfr/index.html.

© 2006 by Taylor & Francis Group, LLC

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Handbook of Lubrication and Tribology

[9] 21 CFR 178.3570 — Lubricants with Incidental Food Contact. Retrieved November 2004 at http://www.gpoaccess.gov/cfr/index.html. [10] Food-grade lubricants, Machinery Lubrication Seminar, Noria Corporation, 209, 2004. [11] Mackwood, W. and Muir, R., Calcium sulfonate complex grease: the next generation food machinery grease. NLGI Spokesman, 17, 2003. [12] Food Processing Industry. Lubricant Selector Guide. Lubrication Engineers, Inc. (no date) [13] Lesinski, D.J. and Raab, M.J., Brand insurance: using the right food lubricant to protect your company. Machinery Lubrication Magazine, 54, 2003. [14] Major Religions of the World Ranked by Number of Adherents, retrieved from Adherents website December 2004 at http://www.adherents.com/Religions_By_Adherents.html, last updated September 2002.

© 2006 by Taylor & Francis Group, LLC

26 Aviation Industry 26.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2 Lubrication of Aviation Piston Engines . . . . . . . . . . . . . .

26-1 26-1

Types of Engines • Wet Sump Engines • Dry Sump Engines • Pressure Section • Scavenging Section • External Lubrication System • Supply Tanks • Oil Temperature Control Devices • Oil Pressure and Temperature Gauges • Types of Piston Engine Oil • Grades of Piston Engine Oil • Important Oil Properties • Oil Drain Intervals

26.3 Lubrication of Aviation Turbine Engines . . . . . . . . . . . .

H.A. Poitz and R.E. Yungk Air BP Lubricants

26-10

Turbine Engine Lubrication Systems • Turbine Engine Oils • Classification of Turbine Engine Oils • Important Oil Properties • Spectrographic Analysis (SOAP)

26.4 Airframe Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

26-13 26-15

26.1 Introduction In 2003 the U.S. aviation industry consisted of approximately 219,000 active aircraft including those operated by the airlines, general aviation, helicopters, and both piston engine powered and turbine powered aircraft. The majority of this number (211,000) is piston engine powered. The amount of lubricants consumed annually by this aviation population consists of about 8 million gal of piston engine oil. A similar volume of turbine oil is also consumed, but some of this is used in marine and power generation equipment. The General Aviation and Regional/Commuter segments of the industry are expected to enjoy the most future growth. Most of this growth will be turbine engine powered. Factors for this turbine powered aircraft growth are the increases in fractional jet ownership and regional jets providing point-to-point service [1].

26.2 Lubrication of Aviation Piston Engines 26.2.1 Types of Engines There are basically two types of piston engines involved in general aviation: radial engines (Figure 26.1) and “so-called” flat engines or horizontally opposed engines (Figure 26.2). Most radial engines have been manufactured by Pratt & Whitney or Curtiss-Wright. The bulk of general aviation is powered by flat engines manufactured by either Textron Lycoming or Teledyne Continental Motors. 26-1

© 2006 by Taylor & Francis Group, LLC

26-2

FIGURE 26.1

Handbook of Lubrication and Tribology

Pratt & Whitney R-1830 Twin Wasp — a 1200 HP, 14 cylinder radial engine.

Various subtypes of these flat engines can be indicated by a letter prefix such as the letter) meaning opposed cylinders, VO — vertical opposed such as in helicopter use, IO — fuel injected cylinders opposed, TIO — turbo charged cylinders opposed, and R — radial cylinders not opposed. The subscript such as O-360 indicates not only that the cylinders are opposed, but also that the total displacement of the cylinders is 360 in.3 .

26.2.2 Wet Sump Engines Most low output engines used in the so-called “light” or “private” plane class employ the “wet sump” type of crankcase lubrication quite similar to that used in automotive. As indicated in Figure 26.3, the crankcase is the oil reservoir in the wet sump system. From this source the oil passes through an internal strainer into a gear pump which forces the oil under pressure into a passage leading to the propeller end of the camshaft. From here the oil flows through the hollow camshaft to the cam bearings. Each cam bearing is provided with an annular groove from which the oil passes through a drilled passage in the crankcase webs to the main bearings. Oil is introduced into the hollow crankshaft through the main bearings and is forced through passages in the crankshaft throws to the connecting rod bearings. Surplus oil escaping from the connecting rod bearings is thrown by centrifugal force onto adjacent parts where it serves to lubricate cylinder walls, piston rings, and pins as well as to cool the pistons as it strikes the underside of the piston heads. In some larger engines actual oil jets impinge on the under side of piston heads to assist in cooling this area and lubricating cylinder walls.

26.2.3 Dry Sump Engines The internal lubricating system of larger aircraft engines and especially radial engines is of the somewhat more complex dry sump type (Figure 26.4). Such systems can be discussed with greater simplicity by

© 2006 by Taylor & Francis Group, LLC

Aviation Industry

FIGURE 26.2

Lycoming AEIO-540 — a 300 HP, opposed 6 cylinder engine.

26-3

© 2006 by Taylor & Francis Group, LLC

26-4

Handbook of Lubrication and Tribology

Breather port Bearings Oil separator

Air

Engine lubrication points Pressure screen

Camshaft, rocker arms Piston rings Push rods, sockets

Oil pump Suction screen

Oil

FIGURE 26.3

Typical wet sump lubrication system on an aircraft engine.

Breather line

Oil separator Breather port

Engine lubrication points

Scavenge pump

Air

Oil cooler Oil filter

Oil tank Lubrication pump

Oil scavenge points

FIGURE 26.4

Oil

Typical dry sump lubrication system on an aircraft engine.

dividing them into pressure section, a scavenging section, and an external system. Pressure section includes the pressure pump, relief valve, check valve, strainers and/or filters, pressure gauge, and the necessary interconnecting pipes and passages. The scavenging section includes the collecting sumps, scavenging pumps, pump screens, and the necessary piping for interconnecting the various units. Finally, the external system includes the oil tanks, controls, and instrumentation.

26.2.4 Pressure Section In many modern engines, the pressure and scavenging pumps may be contained in the same housing and driven by the same shaft. The pressure pump inlet is connected in such a manner as to permit the pump to supply oil as long as any remains in the supply tank. The capacity of the pump is sufficient to furnish a surplus volume of oil at ample pressure to all pressure lubricated parts. Surplus oil from these parts lubricates adjacent parts by the splash system. To prevent excessive rise in pressure, a ball or piston-type relief valve is either built into the pump or connected at some point in the pressure system convenient for adjustment. In inverted or radial engines, where all or part of the cylinders operate in an inverted position, another spring-loaded check valve is often provided to prevent oil from flowing into the engine and accumulating in the combustion chambers of the lower cylinders when the engine is not operating.

© 2006 by Taylor & Francis Group, LLC

Aviation Industry

26-5

Oil strainers and filters are usually used to remove solid contaminants such as carbon, dust, dirt, and metal particles from the oil. The filters are commonly installed on the pressure side of the oil pump and may be any one of several types: non-cleanable depth media element, centrifugal, or stack metal disk. It is to be emphasized that no currently known oil filter will successfully remove or completely neutralize the various forms of liquid and oil soluble contaminants resulting from the deterioration of the lubricant itself. One important feature of many aviation engine lubrication systems is an “anti-sludge” tube to prevent the accumulation of sludge in the crankshaft throws from clogging the hole in the crankshaft through which the oil is supplied to the bearing. The anti-sludge tube is simply a small tube connected to the oil hole in the journal and projecting vertically into the large oil reservoir within the hollow crankshaft. The high speed at which the crankshaft rotates exerts considerable centrifugal force on the oil passing through the hollow shaft. As a result, sludge, which might otherwise stop the flow of oil through the oil hole to the bearing, is effectively thrown out and accumulates harmlessly in the cavity inside the crank pin. Drain plugs, located on the “cheeks” of the crankshaft throw, provide openings at both ends of the anti-sludge chamber for cleaning during overhaul. In addition to supplying oil to the engine, the internal lubrication system also provides lubrication for accessories such as the distribution shaft, tachometer drive, vacuum pump drive, supercharger mechanism, and propeller-reduction gears. Furthermore, where controllable pitch or contact speed hydraulic propellers are used, engine oil must be supplied both for their lubrication and their hydraulic actuation.

26.2.5 Scavenging Section Some engines of the larger sizes are provided with as many as three sumps and an equal number of scavenging pumps. In engines such as the twin row radial type, one sump is usually located at a low point in the nose section, another between the rocker arm boxes of the lower cylinder, and the third or main sump at the lowest point in the crankcase. These various sumps are provided to collect oil readily from the different positions in the engine regardless of its attitude. For example, in a “nose down” position most of the oil will flow to and must be scavenged from the main or engine sump, while some also must be collected and scavenged from the nose section. Meanwhile, excess oil from the rocker arm boxes is collected in the sump located between the rocker arm boxes of the lowest cylinder. In a single row radial engine usually one engine sump located at the lowest point in the crankcase collects all the return oil. In connection with the scavenging system, Figure 26.5 shows how excess oil is prevented from flowing into the lower cylinders of a radial engine. Sumps are always provided with drain plugs to permit complete drainage during oil change. The drain plugs are often provided with small magnets which collect metallic particles resulting from wear or breakage. Wire mesh screens may also be located in the lines. The scavenging pump of an aviation engine is very similar in construction to the pressure pump although it is usually considerably larger to keep the sump relatively free from oil under all conditions. The scavenging pump must first lift the oil from the sump and then force the oil through the restricted passages of the oil cooling unit to the top of the oil supply tank. Scavenging or return oil lines from the sump to the scavenging pump may be either externally located or built integrally with the crankcase section. A breather is provided on the upper part of the crankcase to relieve internal pressure resulting from high-speed piston action, blowby, and the high temperatures prevalent in aircraft engines. Such breathers are so constructed as to relieve the pressure without excessive oil loss. In some engines the breather unit may include a gravity operated valve which remains open when the engine is in a normal position and closes when the engine is inverted to prevent oil loss.

26.2.6 External Lubrication System For purposes of clarity the external may be broken down into three units: the supply tank or tanks and their related parts; the oil temperature controlling units; and the necessary pressure gauges, temperature gauges, and related acessories.

© 2006 by Taylor & Francis Group, LLC

26-6

Handbook of Lubrication and Tribology

FIGURE 26.5

26.2.7 Supply Tanks These tanks must be capable of withstanding internal test pressures of approximately 5 lb/in.3 without failure or leakage. The tank should be located as close as possible to the engine and in most cases at a sufficiently higher level to provide for gravity flow to the engine when it is in normal position. Capacity of the supply tanks will vary according to engine size, and regulations require a certain minimum oil capacity per gallon of fuel capacity. If the engine manufacturer’s specifications for oil capacity exceed the minimum governmental requirements, then the capacity must be no less than that recommended by the manufacturer. Regulations also require that oil tanks be properly vented and provided with an expansion space which cannot be unintentionally filled with oil. In preparation for starting in cold weather, viscosity of the oil may be lowered by means of electrical heaters lowered into the oil tank just prior to starting. Electric current from the hangar circuits may operate these auxiliary heating devices.

26.2.8 Oil Temperature Control Devices One of the most important units for control of oil temperature in the external lubrication system is the oil cooler, a heat exchanger similar to an automobile radiator. Air from the outside is the heat exchanging medium, operating in the same manner as air passing through the radiator of an automotive vehicle (Figure 26.6). Two types of control valves are used to determine whether or not the oil should be bypassed around or directed through the radiator core (1) the thermostatic type which is sensitive to change in temperature

© 2006 by Taylor & Francis Group, LLC

Aviation Industry

26-7

Oil in

Oil out

Air flow

FIGURE 26.6

and (2) the spring loaded or pressure type. When the oil returning from the engine is comparatively cool, the thermostatic valve opens and permits the oil to bypass the cooling surface and flow through the core jacket and then to the top of the oil supply tank. As the oil returning from the engine increases in temperature, the thermostatic valve, by means of a temperature sensitive element closes the bypass passage and requires the oil to flow through the cooling core. Thus by varying the amount of oil passing around and through the cooler, the thermostat maintains the required temperature. The spring loaded valve depends entirely upon the pressure in the oil leading from the scavenging pump. When the oil is cool and its viscosity is high, an excessive pressure is required to force it through the cooling passages. This pressure causes the spring loaded valve to open and allow the oil to pass directly into the tank without cooling. As the temperature of the oil increases due to engine operation, its viscosity decreases and less pressure is required to pump it. The valve then closes resulting in the oil being forced through the cooling core. In general, both the thermostatic and the pressure actuated types are calibrated to require passage of the oil through the cooling core when the returning oil reaches approximately 60◦ C.

26.2.9 Oil Pressure and Temperature Gauges To enable the pilot or flight engineer to monitor operation of the lubrication system, temperature and pressure gauges are provided. The pressure gauge usually indicates oil pressure at or near the main gallery line of the engine. The temperature of the oil is generally determined from thermocouples and is indicated on gauges located in the cockpit control panel.

26.2.10 Types of Piston Engine Oil The majority of aircraft piston engines in operation today were developed on a non-additive petroleumbased oil that has been in existence since the 1920s and is presently covered by SAE J1966 [2], formally Military Specification, MIL-L-6082. Little significant change has occurred in these specifications and usage of this type of oil is still sizable. Very few significant improvements in piston engine oil were made until 1958, when nonmetallic dispersants, antioxidants, and antiwear additives were introduced. This type of oil offered significant

© 2006 by Taylor & Francis Group, LLC

26-8

Handbook of Lubrication and Tribology TABLE 26.1 Military grade 1065 1080 1100 1120

Grades of Piston Engine Oil Viscosity range

Commercial grade

SAE grade

cSt at 100◦ C

SSU at 210◦ F

65 80 100 120

30 40 50 60

10.7–12.4 14.2–16.7 18.7–21.1 23.9–26.1

62–68 75–85 93–103 115–125

improvements in engine cleanliness, reduced wear in oil lubricated areas, and improvement in oil life in high temperature operation. Military Specification MIL-L-22851 was written around this type of oil, but this has been superceded by SAE J1899 [3]. This type now represents over 70% of aircraft piston engine oil sold. Although oils meeting this specification soon became incorrectly known as “detergent” oils, they were never intended to clean up dirty engines. Their correct designation is “dispersant” oils, inasmuch as they have the ability to disperse or keep in suspension products of combustion that blew by the piston rings, as well as products of oil degradation. By keeping these products in dispersed condition, they were removed from the engine when the oil was drained and this led to much cleaner engines. To take advantage of the performance of the newer dispersant oil, some engine models have been type-tested on this oil and restricted to its use. Some new engines have been designed and certified to only use dispersant oils.

26.2.11 Grades of Piston Engine Oil SAE Technical Committee TC-8 is the custodian of the two piston oil specifications used today. J1966 (nondispersant) and J1899 (dispersant) oils both include viscosity grades 65, 80, 100, and 120 as well as multigrade viscosities. The larger the grade number, the more viscous the oil. These commercial grade numbers are the approximate midpoint of the SSU viscosity range at 100◦ C. Early Pratt & Whitney and Wright Aeronautical Specifications, as well as those of the smaller engine manufacturers Avco-Lycoming and Continental Motors, only covered Grades 120 and 100. In later years, the smaller engine manufacturers expanded their specifications to include Grades 80 and 65, primarily for use in colder climates. Present-day designations have dropped the four-digit military grade numbers in favor of a two or three digit grade, Grade 1080 now becoming Grade 80. Table 26.1 shows the historical military grade, commercial grade, SAE grade (automotive oil), and viscosity ranges of the four grades of oil.

26.2.12 Important Oil Properties Lubricating oil in an aircraft engine must serve many functions in temperature ranging from −35◦ C to over 95◦ C, and at elevations from sea level to 6000 m. These functions include lubrication of cylinders, piston rings, valves, gears, and bearings to minimize wear; cooling of the engine hot areas; and performing as hydraulic oil for operation of variable pitch propellers as well as other hydraulically operated mechanisms. Because of these many functions, the following properties are usually defined in the many specifications covering aircraft oil: • Viscosity (ASTM D445) — helps provide a correct lubrication film and correct operating temperatures. Both pour point and viscosity play a major role in the crank-ability of the engine at cold temperatures. • Viscosity Index (ASTM D2270) — the higher the viscosity index, the better the oil resists thinning at the elevated temperature at which an aircraft oil performs. The oil must remain thick enough to provide an effective film between hot metal parts yet must be fluid enough to be readily circulated through the oil system.

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• Pour point (ASTM D97) — discussed previously and relates to the ability to crank the engine in cold weather. • Flash point (ASTM D92) is used to control flammability and volatility of the oil. • Sulfur content by x-ray spectroscopy (ASTM D2273) — sulfur not removed in the refining process can react with moisture and other products of combustion to form compounds highly corrosive to aircraft engine bearings and valve guides. Total sulfur is controlled by specifications and the copper strip corrosion test (ASTM D130) controls corrosivity by sulfur compounds. • Neutralization Number (ASTM D664) — acidic material in an oil may corrode bearings and valve guides and must be controlled. • Ash Content (ASTM D482) — low ash content is also desirable to preclude formation of excessive combustion chamber deposits that might turn into miniature preignition sites and cause extensive piston damage. The limits of ash content usually approach zero for aircraft engine oils. • Effect of an oil on metals (ASTM D130) — to predict the effect of the oil on bronze valve guides and other copper containing components. In the newer ashless dispersant oils the following additional properties are quite important: • Wear properties (ASTM D6709) — measures the ability of the oil to minimize wear as the result of the use of antiwear additives. Testing involves measuring wear on various parts of a small engine in prescribed test. • Antifoam (ASTM D892) — especially important at higher altitudes, because excessive foaming could result in pressure pump starvation and damage as well as foaming out of a tank vent system and partial loss of oil. Foaming is usually controlled by a small addition (5 ppm) of a commercial antifoam additive that modifies surface tension of the oil and permits rapid release of entrapped air. All candidate oils show satisfactory performance (wear, deposit control, oil condition control, dispersancy, etc.) in a 150 h endurance test run on a Textron Lycoming TIO-540-J2BD engine under a test program listed in SAE J1899 and J1966.

26.2.13 Oil Drain Intervals Oil drain intervals for aircraft piston engines are established by the engine manufactures. During the time when the major power sources of larger aircraft used by the airlines and daily utilization was high, many airlines operated their engines full overhaul life with no drains for the lube oil. This was possible because the clearances in the large radial engine were such that a large amount of oil escaped by the piston rings and valve stems and was burned by the exhaust system, accounting for very large oil add amounts and a virtual replacement of used oil with makeup or new oil every 40 to 50 h. Flat engines manufactured by Textron Lycoming or Teledyne Continental Motors are usually drained at 50-h intervals unless equipped with approved full-flow filters and then may be extended to 100-h drain intervals. Again, the maximum authorized drain period is specified by the engine manufacturer in the operator’s handbook. Engine manufacturers sometimes allow extension of oil drain periods under very well regulated conditions. For the average operator, however, it is imperative that they abide by the present oil drain periods as specified in the aircraft operating manual. Some factors that influence oil drain intervals can be itemized as follows: 1. Current engine age In a new engine or one that has very low oil consumption because of tight clearances, the oil remains in the crankcase in usage for a considerably longer time than in an engine that is toward the end of its overhaul, and has worn sufficiently to allow higher oil consumption and thus a much higher rate of new oil makeup. With this in mind, oil should be drained more frequently in a new or tight engine than in an older engine where new oil is added quite frequently. 2. Average monthly usage If an aircraft only flies 4 or 5 h on the weekend and then is idle the remainder of the week, a very severe service condition is established with regard to oil deterioration. The engine lands

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Handbook of Lubrication and Tribology

hot, cools down, and water then condenses in the crankcase and reacts with the combustion chamber blowby products to form highly corrosive acids. If these acids are not burned off by operating the engine again, they will tend to corrode all of the bearings and cause considerable degradation to the oil itself. Conversely, an operator that flies every day for long periods of time, for instance 10 h a day, causes much less severe stress on the oil than the “weekend pilot.” 3. Filtration While filtration was once considered to impose quite a weight penalty, recent studies have indicated that good full-flow filters can double the oil drain interval under normal circumstances. Engine manufacturers now issue service bulletins indicating approved filter kits with many types of oil, which can double the authorized oil drain period. 4. Type of oil The most desirable type of oil available at the present time, with regard to compatibility with filters and resistance to sludge and carbon deposit build-up is the ash-less dispersant oil. This type provides the maximum oil drain intervals permitted by the engine manufacturers.

26.3 Lubrication of Aviation Turbine Engines 26.3.1 Turbine Engine Lubrication Systems Turbine engines have lubrication systems very similar to the piston engine dry sump systems described earlier. Temperature can be much higher with compartment drain temperature in excess of 150◦ C. Because of the high heat loads inherent in a turbine engine due to both radiant heat and internal friction, lube systems employ both air and fuel heat exchanges for temperature control. Bearing compartments and gear box accessory (for fuel pumps, generators, air driven starters, oil pumps, etc.) drive shafts are sealed with carbon face seals, labyrinth seals, or hydraulic seals. Static seals

Overboard discharge Breather pressure valve Deoiler

Rear bearing compartment

Gearbox

Front bearing compartment

Mid bearing compartment

Scavenge pump

Scavenge pump

Scavenge pump

Scavenge pump

Chip detector

Chip detector

Chip detector

Chip detector

Deaerator Oil reservoir Main oil pump

FIGURE 26.7

Main oil filter

Typical lubrication system on a gas turbine engine.

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Oil coolers

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are typically accomplished with orings or gaskets made from fluorocarbon elastomers although some nitrile, silicone, and fluorosilicone seals are employed in older designs. Because of the good thermo-oxidative stability of modern day turbine oils, a reasonable level of consumption (0.2 to 1 lb/h), and high utilization rates found in commercial applications, turbine oil is seldom drained on a periodic basis. The oil remains serviceable by topping off lost oil between overhaul periods.

26.3.2 Turbine Engine Oils Aircraft gas turbines developed prior to 1948 were lubricated satisfactorily with light mineral oils. As power requirement increased and engine cycle temperatures rose, however, mineral oils lacked the necessary high-temperature stability. Severe oxidation of the oil resulted in excessive oil thickening. Oxidation and thermal degradation also initiated a search for a new class of oils which resulted in the development of the synthetic aircraft turbine oils. The synthetics in common aircraft use are ester fluids which have several inherently desirable characteristics. One of these is good response of viscosity to temperature (viscosity index). Another is good thermo-oxidative stability, as evidenced by high bulk fluid temperatures (up to 200◦ C) and thermal decomposition temperatures in transient hot spots (up to 370◦ C). Starting with such base oils, these properties are further enhanced by the addition of antioxidants, metal deactivators, load carrying agents, etc. Esters used as lubricating base oils have three or four ester groups attached to a central carbon atom. The ester groups may contain a variety of carbon chain links ranging from four to eleven, depending upon the selection of the alcohol and acids used as a starting material. With the selection of the proper base oil and the incorporation of the necessary additive systems, fully formulated ester oils have several good properties such as (1) excellent deposit control, (2) good lowtemperature fluidity, (3) good oxidative and thermal stability, (4) good bearing and gear fatigue resistance, (5) good gear load-carrying ability, (6) good corrosion protection, and (7) good compatibility with all engine materials of construction, especially non-metallic seals. Among these, the major feature most important to engine operation is good deposit control. Many engines used in aircraft do not receive full overhauls until 20,000 to 40,000 h have elapsed, thus any deposits formed within the engine may cause lubrication system problems. Such problems are manifested by plugging of oil jets, resulting in oil starvation to bearings and gears; plugging of the breather system, resulting in excessive oil system breather compartment pressures, that result in higher than normal compartment temperature and severe oil leakage; plugging of scavenge systems, resulting in severe oil loss; and carbon shaft seals being rendered inoperative thereby allowing excessive air leakage into the oil system, resulting in excessively rapid oil degradation and deposit formation.

26.3.3 Classification of Turbine Engine Oils Most of the turbine engine oils used today are described in military specifications MIL-PRF-23699 [4] and MIL-PRF-7808 [5]. The oil described by MIL-PRF-23699 is a 5 cSt viscosity grade (at 100◦ C) and serves not only U.S. Army and Navy turbine applications but also most commercial and general aviation turbine applications worldwide. There are now three classes of oil described in MIL-PRF-23699: STD or Standard class, HTS or High Thermal Stability class, and C/I or Corrosion Inhibiting class. The HTS class of oil uses very stable ester basestocks and advanced antioxidant packages to achieve optimized thermo-oxidative stability for the purpose of minimizing deposits that limit time between overhauls (TBO). The C/I class of oil uses a corrosion inhibition package to minimize corrosion in military applications where operating environments are particularly corrosive. MIL-PRF-7808 consists of two viscosity grades, 3 and 4 cSt (at 100◦ C) and is utilized in applications requiring low temperature start capability as low as −54◦ C. These applications include U.S. Air Force fighter aircraft in cold climates as well as commercial aircraft auxiliary power units (APUs) that endure

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Handbook of Lubrication and Tribology TABLE 26.2

Grades and Classes of Turbine Engine Oil

Specification

Viscosity cSt at 100◦ C

MIL-PRF-7808, Grade 3 MIL-PRF-7808, Grade 4

3 4

MIL-PRF-23699, Class STD MIL-PRF-23699, Class HTS MIL-PRF-23699, Class C/I DOD-PRF-85734

5 5 5 5

Attributes Lowest temperature start capability Improved cleanliness/stability with good low temperature start capability Higher viscosity for improved system durability Best cleanliness/stability Corrosion Inhibition for salt environments High load carrying for helicopter gearboxes

cold temperatures at altitude. The more viscous MIL-PRF-7808, Grade 4 also has higher thermal stability requirements recognizing the need for more stable oils for advanced military aircraft that operate at increasingly hotter temperatures. So-called high load carrying oils are typically used for helicopter transmission gearboxes and some gas turbine systems with highly loaded gear sets. These 5 cSt grade oils are described in DOD-PRF-85734 [6] are similar to the MIL-PRF-23699 oils with an additional Extreme Pressure (EP) additive system that promotes better boundary film lubrication in high load situations. The forum for many turbine oil issues, evaluation method development and future specifications is an SAE Standards Development committee, E-34 (Propulsion Lubrication). E-34 published a specification, AS 5780 [7], for 5 cSt turbine oils used in commercial applications. Original equipment manufacturer (OEM) are using this specification as part of their regulated oil approval activities.

26.3.4 Important Oil Properties In addition to the properties important to piston engine oils, turbine engine oils are evaluated for other lubricant attributes appropriate to their base chemistry, engine materials of construction and engine time between overhauls. Lubricant compatibility (FED-STD-791, Method 3403) evaluates the hot aging miscibility of candidate oils with other oils likely to be encountered in service. This is important given the possible additive package diversity as allowed by performance (not material) specifications. An acid assay (FED-STD-791, Method 3500 (1)) is performed to control the acid distribution used in an oil formulation’s ester basestock. This ensures the oil chemistry is controlled on a batch-to-batch basis similar to the originally qualified formulation. Elastomer compatibility (e.g., Def Stan 05-50 (Part 61) Method 22) measures swell and deterioration of seal materials in contact with hot oil. Elastomers of interest include fluorocarbon, perfluorocarbon, fluorosilicone, silicone, and nitrile. This becomes important because the ester basestocks and antioxidant packages tend to be aggressive toward fluorocarbon, while extreme pressure additives are aggressive toward silicone and fluorosilicone. Oxidation and Corrosion Stability (FED-STD-791, Method 5308 mod 1 or ASTM D4636) evaluates bulk oil stability at temperatures up to 218◦ C as well as any breakdown products consequential corrosive impact on representative metallurgies. Thermal Stability and Corrosivity (FED-STD-791, Method 3411) has proven an effective quality control method for detecting undesirable contaminations from other non-aviation ester based products. Deposit control is measured in a full scale, heated bearing rig (FED-STD-791, Method 3410, severity rating of 1.5). This test is run for 100 h (HTS oils require 200 h) with evaluation of oil condition control and a cleanliness demerit rating. Emerging subscale tests are being developed and evaluated to determine control of liquid phase deposits (SAE ARP 5996 to simulate oil pressure pipes and jets), mixed phase deposits (GE Alcor High Temperature Deposition Test to simulate scavenge system pipes), and vapor phase deposits (U.S. Navy test to simulate

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breather system pipes). This activity is indicative of the importance of controlling deposits in engines with ever increasing time between overhauls. Load carrying or boundary film lubricating ability is measured by the Ryder Gear test (FED-STD-791, Method 6508 and/or SAE AIR 4978 Appendix E). This becomes especially import for the DOD-PRF-85734 oils that are used in highly loaded helicopter transmission gearboxes. Hydrolytic stability (Def Stan 05-50 (part 61) Method 6) is important to storage stability and the stability of oils in capped (not breathed) lubrication systems often found in engine accessories such as generators and air starters. Hydrolysis is a degradation chemical reaction between an ester and water at elevated temperatures that reverts the ester to acid and alcohol destroying oil properties. Corrosion Inhibition in MIL-PRF-23699, Class C/I oils is controlled by SAE ARP 4249 ball corrosion testing. Trace metals are controlled because new oil needs to provide a baseline for spectrographic oil analysis program diagnostic testing discussed later.

26.3.5 Spectrographic Analysis (SOAP) The parts per million (ppm) of several metals (e.g., Fe, Ag, Cr, Al, Mg, Ti, Mo, V, etc.) in a sample of used oil is usually determined by exposure of a sample to an emission spectrometer. Spectrographic oil analysis was first used in the 1940s by the railroads to determine bearing wear in diesel engines, and was very successful. The military studied its use in predicting power plant failure in aircraft engines and the practice is now in widespread use. Many analytical laboratories offer such service to airlines and operators of general aviation. Provided the proper techniques are utilized in obtaining oil samples and sufficient background information is provided to the analytical laboratory on the type of engine and oil history, spectrometric analysis is a very useful tool in determining the condition of the oil. Caution should be observed in the use of one random sample of oil to assess any engine condition. Each engine has its own normal wear pattern: only after observing such a pattern over several oil changes in many hours of engine operation will significant deviations from this pattern become meaningful. Most operators use a combination of engine parameter trending (oil consumption, temperature, pressure, etc.), periodic examination of magnetic chip collector, examination of oil filter/screen for wear debris, and spectrometric oil analysis to judge the condition of their engines. One cardinal rule is to immediately obtain a second sample of oil for analysis if an alarm is sent back from the laboratory as to potential engine problems as observed by excessive wear metals in the spectrometric analysis. Spectrometric oil analysis of oils from turbine-powered aircraft is somewhat more meaningful than similar analysis of oils from piston-powered aircraft because of the absence of combustion by-products in the turbine oil. These by-products may mask determinations of some wear metals. Magnetic chip collectors incorporated into the scavenge system also serve as a valuable diagnostic compliment for engine condition monitoring. These chip collectors are inspected by mechanics at regular intervals. The debris collected can be evaluated with scanning electron microscopes equipped with energy dispersive spectroscopy (SEM/EDS) to ascertain the material type and wear mechanism by which the debris was generated. Advanced lubrication system designs are beginning to incorporate oil and debris monitors that are integrated into the engine’s control system providing a more comprehensive Prognostic Health Monitoring (PHM) system.

26.4 Airframe Lubrication Selection of proper lubricants, frequency of lubrication, and points requiring lubrication are the responsibility of the airframe manufacturer. This is true for all aircraft, whether a Piper Cub, or a Boeing 747 aircraft. A subcomponent manufacturer, for example an engine manufacturer or a flap gearbox manufacturer,

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250

CARBURETOR & CABIN HEAT GUIDE

100

ELEVATOR PULLEYS THROTTLE LEVERS

100

LUBRICANT HOURS IDLER PULLEYS 250 (SEE CAUTION 3)

100 RUDDER HINGES

FLAP PULLEY STABILIZER ADJUSTMENT PULLEY (SEE CAUTION 3) 250

RUDDER & 100 ELEVATOR HORNS

BRAKE MASTER CYLINDERS

ELEVATOR HINGES 100 LEFT & RIGHT

50

CONTROL STICK BEARINGS TORQUE TUBE BEARINGS 100

ELEVATOR PULLEY

ENGINE OIL SUMP DRAIN AND REFILL (SEE NOTE 6, PAGE 3)

100 TAIL WHEEL BEARING

100

50

26-14

LUBRICATIONCHART

HOURS LUBRICANT FLAP PULLEYS

50

TAIL WHEEL SWIVEL

STABILIZER ADJUSTMENT 100 MECHANISM (SEE CAUTION 3)

ENGINE

100 GREASE FITTING ELEVATOR & STABILIZER PULLEYS, FLAP HINGE 100 BEARINGS, FLAP CRANK & PUSH ROD BEARINGS (SEE NOTE 4)

RUDDER PEDAL BEARINGS BRAKE PEDAL BEARINGS 100 FLAP HANDLE BEARINGS FLAP HANDLE RATCHET 100

AILERON HINGE BEARINGS AILERON HORN 100 AILERON PULLEYS LEFT & RIGHT (SEE NOTE 4)

SHOCK STRUT PIVOTS

100

LANDING GEAR WHEEL BEARINGS

100

AILERON PULLEY 100 RUDDER PULLEY LEFT & RIGHT

GEAR HINGES 100 LANDING LEFT & RIGHT

LEGEND MIL-G-23827

NOTES 1. CARBURETOR AIR FILTER — CLEAN PER MANUFACTURER’S INSTRUCTIONS ON FILTER BOX OR INSTRUCTIONS IN OWNER’S HANDBOOK. (UNDER ABNORMAL CONDITIONS, FILTER REQUIRES CLEANING MORE FREQUENTLY. REPLACE AS REQUIRED.) 2. LUBRICATION POINTS — WIPE ALL LUBRICATION POINTS CLEAN OF OLD GREASE, OIL, DIRT, ETC. BEFORE RELUBRICATING. 3. WHEEL BEARING REQUIRES CLEANING AND REPACKING AFTER EXPOSURE TO ABNORMAL QUANTITY OF WATER. 4. AILERON AND FLAP HINGES-HINGE BLOCKS WITH LUBRICATION HOLES IN THEIR UNDERSIDE MAY BE PRESSURE LUBRICATED WITH GREASE MIL-23827

MIL-L-7870 MIL-L-3545 MIL-H-5606

ENGINE

GREASE, AIRCRAFT AND INSTRUMENT, GEAR AND ACTUATOR SCREW OIL-GENERAL PURPOSE LOW TEMP. LUBRICATION GREASE-LUBRICATION HIGH TEMPERATURE HYDRAULIC FLUID (RED)

SAE 50 ABOVE 60 F AIR TEMP SAE 40 BETWEEN 30 F AND 90 F AIR TEMP LYCOMING 0-290-D2 & 0-320 ENG. SAE 30 BETWEEN 0 F AND 70 FAIR TEMP SAE 20 BELOW 10 F AIR TEMP SAE 20 BELOW 32 F AIR TEMP SAE 40 ABOVE 32 F AIR TEMP CONTINENTAL C90 ENGINE SEE LYCOMING SERVICE INSTRUCTIONS NO. 1014 FOR USE OF DETERGENT OIL

FIGURE 26.8 © 2006 by Taylor & Francis Group, LLC

CAUTIONS 1. DO NOT USE A HYDRAULIC FLUID WITH A CASTER OIL OR ESTER BASE. 2. DO NOT APPLY LUBRICANT TO RUBBER PARTS 3. TRIM CABLES — UNDER NO CIRCUMSTANCES SHOULD THE TRIM CABLES FROM THE COCKPIT TO THE REAR OF THE FUSELAGE BE LUBRICATED. (TO PREVENT SLIPPAGE) 4. CONTROL CABLES — WIPE CLEAN AT REGULAR INTERVALS BUT DO NOT LUBRICATE. UNDER SALT CONDITIONS OCCAOCCASIONAL LUBRICATION WITH MIL-L-7870 IS RECOMMENDED.

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50

CARBURETOR AIR FILTER (SEE NOTE 1)

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may also specify a certain oil or grease for their particular unit, but final responsibility for the correct lubricant rests with the airframe manufacturer. Each aircraft has a lubrication chart provided by the airframe manufacturer (e.g., Figure 26.8) [8] that clearly specifies the type of lubricant (usually by a military specification), the point of lubrication, and the frequency of lubrication. In the case of aircraft, this figure is usually expressed in hours of operation instead of miles. If an aircraft is operated infrequently, that is less than 20 h/month, the lubrication intervals are normally shortened by half of what is shown in the maintenance manual. By specifying lubricants by military specification (consensus OEM specifications are under development) the airframe manufacturer does not restrict the buyer to one brand name product and widens the availability of the lubricant geographically. One brand-name product is occasionally recommended on an exclusive basis because of superior performance, but airframe manufacturers try to avoid this situation. The industry forum for developing both OEM based specifications and test methods are SAE AMS M Aviation Greases committee.

References [1] FAA Aerospace Forecasts, Fiscal Years 2004–2015, U.S. Department of Transportation, Federal Aviation Administration, Office of Aviation Policy and Plans, March 2004. [2] SAE Specification J1966, Lubricating Oils, Aircraft Piston Engine (Non-Dispersant Mineral Oil), 2005. [3] SAE Specification J1899, Lubricating Oil, Aircraft Piston Engine (Ashless Dispersant), 2005. [4] MIL-PRF-23699 Performance Specification, Lubricating Oil, Aircraft Turbine Engine, Synthetic Base, NATO Code Number O-156, 1997. [5] MIL-PRF-7808 Performance Specification, Lubricating Oil, Aircraft Turbine Engine, Synthetic Base, 1997. [6] DOD-PRF-85734 Performance Specification, Lubricating Oil, Helicopter Transmission System, Synthetic Base, 2004. [7] SAE Aerospace Standard AS5780, “Core Requirement Specification for Aircraft Gas Turbine Engine Lubricants,” 2005. [8] Lubrication Chart for the Piper Super Cub Aircraft provided by Piper Aircraft Corporation, 2002.

© 2006 by Taylor & Francis Group, LLC

27 Lubrication for Space Applications 27.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.2 Lubrication Regimes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27-1 27-2

Hydrodynamic, EHL, Mixed, and Boundary Lubrication • Factors Influencing Boundary Film Formation

27.3 Liquid Lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27-3

Types of Liquid Lubricants • Liquid Lubricant Properties

27.4 Greases and Solid Lubricants . . . . . . . . . . . . . . . . . . . . . . . . .

27-13

Greases • Solid Lubricants

27.5 Mechanism Components and Re-Lubrication Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27-17

Spacecraft Components • Re-Lubrication Mechanisms

27.6 Lubricant Testing and Analysis . . . . . . . . . . . . . . . . . . . . . . .

27-20

Types of Testing • Accelerated Testing • Facilities for Space Lubricant Testing

William R. Jones and Mark J. Jansen NASA Glenn Research Center Tribology and Surface Science Branch

27.7 Lubricant Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27-29

Relative Life and Wear Characteristics • General Mechanism Effects • Cleaning and Surface Preparation

27.8 Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27-33 27-33

27.1 Introduction Space tribology is a subset of the lubrication field dealing with reliable satellite and spacecraft performance. It encompasses the entire gamut of tribologic regimes, including elastohydrodynamic lubrication (EHL), parched EHL, transient EHL, boundary lubrication, and mixed lubrication. Historically, choices of space mechanism lubricants were based on space heritage rather than on the latest technology or best available materials. With the limited mission lives and minimal duty cycles of the early space program, this strategy was highly successful. As missions extended, other spacecraft components such as electronics, batteries, and computers, failed before lubricated mechanisms [1]. However, during the 1980s and 1990s, these ancillary components vastly improved and tribologic systems have become one main factor limiting spacecraft reliability and performance. Although tribologic components represent only a small fraction of the spacecraft’s cost, they are often single point failures that cripple or debilitate expensive spacecraft.

27-1

© 2006 by Taylor & Francis Group, LLC

27-2

Handbook of Lubrication and Tribology h~0.0025 – 1.25 mm

h~0.0025 mm

h~0.0025 – 0.025 mm h >0.25 mm Hydrodynamic

Elastohydrodynamic

Mixed

Boundary

Cofficient of friction

.150

.001 Viscosity × Velocity , ZN Load P

FIGURE 27.1 Coefficient of friction as a function of viscosity–velocity–load parameter. (From Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.)

The Galileo spacecraft is a classic example of single point tribologic failure affecting the entire mission. The Space Shuttle Atlantis launched Galileo in 1989, starting a six-year journey to Jupiter. A high gain antenna, used to transmit control and telemetry data to earth, was one of the craft’s most important components. The umbrella shaped antenna was stowed closed behind a sun shield. In 1991, after the craft’s pass near the sun, the antenna was deployed, but it only opened partially. Engineers concluded three of the 18 ribs in the antenna’s umbrella-like structure were stuck in place. Ground-based tests [2] showed titanium alignment pins, which had been lubricated with a bonded dry film lubricant, had galled and subsequently seized due to lack of lubricant, causing the antenna failure. In this case, engineers salvaged the mission using the low gain antenna combined with data transmission advancements. This chapter will discuss basic lubrication ideas that play a major role in space mechanisms, space lubricant types, details of the most common space lubricants, mechanism components, testing and test facilities, and factors affecting lubricant selections.

27.2 Lubrication Regimes Lubrication separates surfaces in relative motion by interposing a third body that has low shear resistance, thus preventing serious surface damage or wear. The third body can be a variety of different materials including adsorbed gases, reaction films, and liquid or solid lubricants.

27.2.1 Hydrodynamic, EHL, Mixed, and Boundary Lubrication Depending on the third body type and thickness, several different lubrication regimes can be identified and are depicted in the Stribeck curve (Figure 27.1). Stribeck performed a series of journal bearing experiments in the early 1900s measuring friction coefficients as a function of load, speed, and temperature [3]. Later, Hersey performed similar experiments and devised a plotting format based on a dimensionless parameter, ZN/P [4]. The Stribeck/Hersey curve plots friction coefficient as a function of viscosity (Z), velocity (N), and load (P).

© 2006 by Taylor & Francis Group, LLC

Lubrication for Space Applications

27-3

When the ZN/P value is high, surfaces are completely separated by a thick (>0.25 µm) lubricant film and occur at high speeds, high viscosities, and/or low loads. In this region, termed hydrodynamic lubrication, lubricant rheology determines the friction. As the ZN/P parameter decreases, the lubrication regime changes from hydrodynamic to elastohydrodynamic, then to mixed, and finally to boundary. The EHL regime occurs in nonconformal, concentrated contacts where high loads cause surfaces to elastically deform and pressure–viscosity effects to occur in the lubricant. Film thickness in this regime ranges from 0.025 to 1.250 µm. As ZN/P continues to decrease, film thickness also decreases and surface interactions start taking place. This regime, where both surface interactions and fluid film effects occur, is referred to as the mixed regime. Finally, at low ZN/P values, the boundary lubrication regime is entered, where surface interactions are the primary factor. The boundary lubrication regime is a highly complex arena involving metallurgy, surface topography, physical and chemical adsorption, corrosion, catalysis, and reaction kinetics [5, 6]. Formation of protective surface films, which minimize wear and surface damage, is the regime’s most important characteristic. Typically, space mechanisms operate in the EHL, mixed, or boundary lubrication regimes, with the boundary lubrication regime being the most severe.

27.2.2 Factors Influencing Boundary Film Formation Both lubricant and bearing surface chemistry govern film formation. Additional environmental factors, such as temperature, also influence the lubricant’s film forming ability. The lubricant’s physical properties determine the film’s effectiveness at minimizing wear. Properties affecting film formation include shear strength, thickness, surface adhesion, film cohesion, melting point or decomposition temperature, and bulk lubricant solubility. 27.2.2.1 Starved EHL An EHL subdivision, starved EHL, describes the situation occurring in ball bearings having a restricted oil supply, where pressure build-up in the contact inlet region is inhibited, resulting in a thinner film thickness than calculated by classical EHL theory [7, 8]. Starvation theory was first described by Wedeven [9] in the early 1970s. 27.2.2.2 Parched EHL In many space mechanisms, instrument bearings are lubricated with a minimal amount of oil. When no free bulk oil is available to form a meniscus, starvation theory cannot adequately describe lubricant behavior. Another EHL subdivision, parched elastohydrodynamics, describes this behavior [10,11]. Lubricant films in this regime are so thin that they are immobile outside the Hertzian contact zone. This regime is particularly important for space mechanisms because parched EHL bearings require the least driving torque and have the most precisely defined spin axis, making them an ideal choice for many applications. 27.2.2.3 Transient/Non-Steady State EHL For space mechanisms, transient or nonsteady state behavior is another important EHL area. In this area, load, speed, and contact geometry are not constant over time. Unlike steady state EHL behavior, nonsteady state behavior is not well understood. However, many practical machine elements, including rolling element bearings, gears, cams, and traction drives, operate under nonsteady state conditions. In particular, stepper motors, commonly used in many space mechanisms, operate in this state. This regime has been studied theoretically for line contacts [12–14] and experimentally for point contacts [15,16].

27.3 Liquid Lubricants For space applications, designers use both liquid and solid lubricants. Both have merits and deficiencies, which appear in Table 27.1 [17].

© 2006 by Taylor & Francis Group, LLC

27-4

Handbook of Lubrication and Tribology TABLE 27.1

Relative Merits of Solid and Liquid Space Lubricants

Dry lubricants Negligible vapor pressure Wide operating temperature Negligible surface migration Valid accelerated testing Short life in moist air Debris causes frictional noise Friction speed independent Life determined by lubricant wear Poor thermal characteristics Electrically conductive

Wet lubricants Finite vapor pressure Viscosity, creep, and vapor pressure all temperature dependent Sealing required Invalid accelerated testing Insensitive to air or vacuum Low frictional noise Friction speed dependent Life determined by lubricant degradation High thermal conductance Electrically insulating

Source: Jones, W.R. and Jansen, M.J., “Space Tribology,” NASA TM 2000-203324, March 2000.

27.3.1 Types of Liquid Lubricants In the last three decades, space applications have used many different liquid lubricants, including mineral oils, silicones, esters, and perfluoropolyethers (PFPE). More recently, a synthetic hydrocarbon (Pennzane®) has been replacing many older lubricants. Each lubricant type will be discussed briefly but since the majority of current spacecraft use either a formulated Pennzane® or one of the PFPE lubricants, these two classes will be discussed in detail. 27.3.1.1 Mineral Oils This lubricant class consists of a complex mixture of naturally occurring hydrocarbons with a wide range of molecular weights. Examples include V-78, BP 110, Apiezon C, Andok C (Coray 100) [18], and the SRG series of super refined mineral oils, including KG-80 [19]. The super refined fluids have been highly processed to remove polar impurities, either by hydrogenation or percolation through bauxite. Refining makes them poorer neat lubricants, but greatly improves additive response. While Apiezon C is still commercially available, production of all others was discontinued many years ago. Nevertheless, some companies have stockpiled SRG oils and still use them to lubricate momentum and reaction wheel bearings. SRG oils have an estimated shelf life in excess of 20 years [19]. 27.3.1.2 Esters Esters, which are available in a wide viscosity range, are inherently good boundary lubricants. In the 1970s, British Petroleum developed a triester base lubricant which was laboratory tested but whose production stopped before it flew in space. Another ester, NPT-4 (neopentylpolyol ester), has been used in the past, but is no longer produced. Nye Lubricants also markets a series of low volatility neopentylpolyol esters (UC4 , UC7 , and UC9 ). 27.3.1.3 Silicones This fluid class was used early in the space program but silicones are poor boundary lubricants for steel on steel systems. Boundary lubrication comparisons of this fluid with a PFPE and a PAO have been reported [20] and Figure 27.2 shows relative lifetimes. Silicone performed poorly, degrading into an abrasive, polymerized product. Versilube F-50, a chloroarylalkylsiloxane, is an early example of this lubricant class. 27.3.1.4 Synthetic Hydrocarbons Two synthetic hydrocarbon groups are available today: polyalphaolefins (PAO) and multiply alkylated cyclopentanes (MACs).

© 2006 by Taylor & Francis Group, LLC

Lubrication for Space Applications

27-5

16 14

Relative life

12 10 8 6 4 2 0 Chloroarylalkylsiloxane

PFPE

PAO

Lubricant type

FIGURE 27.2 Screening test results (scanner and mechanism). (From Didziulis et al., “Lubrication,” NASA/TP-1999206988, 1999.) TABLE 27.2 Typical Properties for Three Commercial Polyalphaolefins

Viscosity at: 210◦ F, SUS 210◦ F, cSt 100◦ F, SUS 100◦ F, cSt 0◦ F, cSt Flash point Pour point Evaporation 6 12 h at 350◦ F Specific gravity at 25◦ C

Oil 132

Oil 182

39 3.9 92 18.7 350 440◦ F −85◦ F 2.2% 0.828

62.5 10.9 348 75.0 2700 465◦ F −60◦ F 2.0% 0.842

Oil 186 79.5 15.4 552 119 7600 480◦ F −55◦ F 1.9% 0.847

27.3.1.4.1 Polyalphaolefin (PAO) Polyalphaolefins are made by the oligomerization of linear α-olefins having six or more carbon atoms [21]. Nye Lubricants markets a number of PAOs for space applications and properties for three commercial PAOs appear in Table 27.2. A new synthetic hydrocarbon based on PAO chemistry has been developed [22]. 27.3.1.4.2 Multiply Alkylated Cyclopentanes Multiply alkylated cyclopentanes (MACs) make up the second hydrocarbon class. These materials are synthesized by reacting cyclopentadiene with various alcohols in the presence of a strong base [23,24]. The reaction products are hydrogenated to produce the final product, which is a mixture of di-, tri-, tetra-, or penta- alkylated cyclopentanes. Varying reaction conditions control the distribution. Originally, only one product, known as Pennzane® SHF-X2000 or Nye Synthetic Oil 2001A, was available for space applications, and is primarily the tri-2-octyldodecyl substituted cyclopentane [23]. However, Pennzane® SHF-X1000, a lower viscosity but higher volatility version, is now available [25]. SHF-X1000 is primarily a di-substituted cyclopentane. A variety of formulated versions for both oils are also available. Properties of SHF-X1000 and SHF-X2000 appear in Table 27.3. Recent experience with SHF-X2000 appears in Carré et al. [26]. A six-year life test of a CERES elevation bearing assembly using a Pennzane®/lead naphthenate formulation yielded excellent results [27]. Additional life test data for another Pennzane®/lead naphthenate formulation for the MODIS instrument appears in VanDyk et al. [28].

© 2006 by Taylor & Francis Group, LLC

27-6

Handbook of Lubrication and Tribology TABLE 27.3

Typical Properties of Two Pennzane® Fluids

Property Viscosity at 100◦ C (cSt) Viscosity at 40◦ C (cSt) Viscosity at −40◦ C (cSt) Viscosity index Flash point (◦ C) Fire point (◦ C) Pour point (◦ C) Specific gravity at 25◦ C Coefficient of thermal expansion (cc/cc/◦ C) Total weight loss, 24 h, 125◦ C, 10−5 Torr Refractive index at 20◦ C Vapor pressure at 125◦ C (Torr)

TABLE 27.4

SHF-X1000

SHF-X2000

9.4 60 N/A 131 290 N/A −52 0.85 N/A 1

DC 06 AISI 304

Source: Roughness data extracted from de Rooij, M.B., Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente, 1998.

Williamson [6], see Equation (28.2). Roughness is represented in this equation by its probability asperity height distribution, with standard deviation σs . Furthermore, it is assumed that summits are equally shaped, that is: spherically with radius β. As a first approximation, it may be assumed that elastic contact occurs when ψ < 0.6 and plastic contact when ψ > 1. The intermediate phase between 0.6 and 1 corresponds with the mixed elastic–plastic regime. ψ=

E
· 2 · Hsoft




σs β

(28.2)

Applying Equation (28.2) and using typical sheet material values from Table 28.2, it is found that the quotient of E
over H is about 64 for stainless steel sheet and 128 for deep draw steel DC 06. Since the square root is at least about 0.05 (for stainless steel), plastic contact is likely to occur at the sheet asperity level. As run-in does not occur for the sheet material, a transition to an elastic regime is not likely to occur. The environment of the tribological contact is typically humid air at room temperature, although exceptions exist, that is, in hydro-forming and hot metal gas-forming applications.

28.2.2 Operational Variables The operating variables for tribological contacts in metal forming depend heavily on the actual application. Furthermore, differences exist between locations at the tool surface. The normal pressure at the blank holder–work piece interface, for example, is much lower than the normal pressure at the die radius, for a deep drawing setup, see Figure 28.5. A general overview of possible contact conditions in metal forming applications is given in Table 28.3. The presented values for the normal pressure can be much higher when considering critical contact situations at drawing radii or during forming of high(er) strength steels. As an example, the sliding velocity and normal pressure for a simple air bending set up are calculated. Tooling will increase in temperature from room temperature to about 40 to 70◦ C, as a result of the forming action. In some cases, forming is done at elevated temperature to make use of specific forming behavior of sheet materials. Indicative temperatures for warm forming and hot forming are 300–600◦ C and 800–1000◦ C, respectively. Let us take the example of air bending. Two tribological contact situations can be identified during the air bending process: the contact between the punch nose and the sheet and the contact between the sheet material and the die shoulder radius. The latter contact is of high importance since failure will induce fluctuating friction forces, which in turn will influence the amount of spring back that is generated after lifting the punch. Furthermore, galling can be initiated in this sliding contact. The operational conditions for the die shoulder–sheet contact has been estimated by ter Haar [7]. First, a time dependent function s(t ) based on the geometry of the process, see Figure 28.6, is derived, which describes the position of the contact on the sheet, relative to the original position for punch displacement, d = 0. s(t ) = lPD (t ) + r · α(t ) + b − a

© 2006 by Taylor & Francis Group, LLC

(28.3)

28-8

Handbook of Lubrication and Tribology F

FN

FN FN F FZ

F FZ FZ

FZ

Blank holder

Die

Forming part

FIGURE 28.5 Tribological contacts for a deep drawing process. Adapted from Netsch, T., Methode zur Emittlung von Reibmodellen fur die Blechumformung, PhD Thesis, Darmstadt University of Technology, Germany, Shaker-Verlag Aachen, Band 41, 1998.

TABLE 28.3

Operational Conditions in Metal Forming Processes Processes

Conditions Contact pressure ratio, p/Y a Contact pressure, p [MPa] Velocity, v [m/sec]

Sheet forming

Drawing ironing

Rolling rotary forming

Forging extrusion

0.1–1 1–100 10−3 –10−1

1–3 100–1000 10−2 –102

1–3 100–1000 10−2 –102

2–5 100–3000 10−3 –10−1

a Y : yield stress of the sheet material.

Source: Adapted from Wang, Z.G., J. Mater. Process. Technol., 151, 223–227, 2004.

with lPD (t ) =

 (a − b)2 + (d(t ) − r − R)2 − (r + R)2

 α(t ) = arc sin  α(t ) = arc sin

r +R l ∗ (t ) r +R l ∗ (t )

l ∗ (t ) =

© 2006 by Taylor & Francis Group, LLC



 − arc cos



 + arc cos

a−b l ∗ (t ) a−b l ∗ (t )

(28.4)

 r + R > d(t ) 

(28.5) r + R ≤ d(t )

 (a − b)2 + (d(t ) − r − R)2

(28.6)

Friction and Wear in Lubricated SMF Processes

28-9

d

Sheet

Punch a

Punch

r

Sheet

Lpd 2b

a

R

Die

Die

FIGURE 28.6 Two stages of air bending. Adapted from: ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.

The time derivative of s(t ) is the velocity of the displacement of the contact point, v(t ): v(t ) =

d [s(t )] dt

(28.7)

The sliding velocity can now be calculated by assuming a certain function for the punch displacement with time. The bending moment per unit width, M , needed for bending the sheet around the punch nose is estimated by ter Haar [7] based on the work of de Vin [8]: 2 M = r 2E
3



E
C∗



3/(n+1)

(St /2)n+2 − r n+2 (E
/C ∗ )(n+2)/(n−1) (n + 2)r n

+ 2C

∗

C∗ =

 (n+1)/2 4 ·C 3

 (28.8)

with (28.9)

This bending moment has to be supplied by the force FN at the die shoulder radius at a distance lPD . Hence: FN (t ) =

B·M lPD (t )

(28.10)

For reasons of simplicity, it is assumed that the contact between the die shoulder and the sheet material remains elastic during the forming process. The maximum contact pressure is found by applying the Hertzian theory for a strip contact:
pmax =

FN /B · E
π · R∗

(28.11)

Here, B represents the length of the line contact, E
the combined modulus of elasticity (Equation [28.1]), and R ∗ the combined radius (Equation [28.13]). R∗ =

© 2006 by Taylor & Francis Group, LLC

R1 R2 R1 + R 2

(28.12)

28-10

Handbook of Lubrication and Tribology TABLE 28.4 Input for Calculation of Contact Pressure and Sliding Velocities During Air Bending Parameter

Value

Unit

Parameter

Value

Unit

dmax B a b r R tstroke

7.5 40 9.67 0.40 1.67 2.00 1

mm mm mm mm mm mm sec

st E1 = E2 υ1 = υ2 C n η —

1 2.1 × 1011 0.33 500 × 106 0.204 0.2 —

mm N/m2 — N/m2 — Nsec/m2 —

Source: Adapted from ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.

5

500 495

4

490

pmax [MPa]

480

2

475 1

470 465

v [mm/sec]

3

485

0

460

pmax [MPa]

455

v [mm/sec]

–1 –2

450 0

0.1

0.2

0.3

0.4

0.5 t [sec]

0.6

0.7

0.8

0.9

1

FIGURE 28.7 Calculated maximum Hertzian contact pressure and the sliding velocity for the die shoulder–sheet contact of the air bending process described Table 28.3.

The calculated values for the maximum contact pressure and sliding velocity as a function of time, for the geometry listed in Table 28.4, are shown in Figure 28.7. Applying the rule of the thumb pmax < 0.6 H with H = 900 GPa, Table 28.2 shows that the assumption of a Hertzian contact is valid for this particular configuration. In general, however, a finite element method calculation is needed to calculate the plastic component of the deformation.

28.2.3 Sheet Materials Sheet metal forming processes are used to produce a variety of products from automotive and aerospace applications to household and kitchen applications. Each product has to meet a specific set of demands with respect to, for example, corrosion resistance, toughness, strength, and visual appearance. Therefore, a range of sheet materials exists that varies in composition, forming behavior and surface quality. A summary of relevant properties for four groups of sheet materials, that is, low carbon steel, stainless steel, aluminum alloys, and copper alloys, is given in Reference 9, see Table 28.5. Low carbon steels represent the largest fraction of sheet material used nowadays in the forming industry. Automotive panels, for example, are typically made of 0.8 mm cold rolled deep drawing steel DC 06, alloyed low carbon steel with high formability. The nominal composition of this sheet material

© 2006 by Taylor & Francis Group, LLC

Friction and Wear in Lubricated SMF Processes TABLE 28.5

28-11

Survey of Representative Sheet Materials

Material

Typical example

Yield stress [MPa]

Tensile strength [MPa]

Brinell hardness [HB]

180 500–620

270–350 570–710

80 —

Carbon steel

DC 06 HSLA Docol 500 YP

Stainless steel

AISI 304/X5 CrNi 18 10 AISI 316/X5 CrNiMo 17 12 2 AISI 409/X2CrTi12 AISI 430/X6Cr17

210 220 220 260

520–720 520–670 380–560 430–630

150–190 150–190 150–180 150–190

Al-alloys

AA 1050A/Al99.5 (O) AA3103/AlMn1 (O) AA 5086/AlMg4 (O) AA 2024/AlCu4Mg1 (T4)

20 30 100 275

65 90 240 425

20 27 65 120

Cu and Cu-alloys

SW-Cu (F22) CuZn 30 (F27) CuSn8 (F33)

max. 140 max. 160 max. 190

220–260 270–350 330–380

40–65 55–85 65–95

Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.

TABLE 28.6 Characteristic 2D Roughness Values, Bulk Hardness, Zinc Layer Thickness, and Hardness, Measured for Galvanized Steel

GI EDT EG GA

Ra [µm]

Rz [µm]

Rq [µm]

Rp [µm]

Rt [µm]

Bulk [HV0.2 ]

Layer [µm]

Layer hardness

0.9 1.2 0.9

4.3 7.0 7.1

1.0 1.5 1.2

1.9 3.6 2.5

5.0 10.1 8.1

81 77 73

14 8 10

46–57 HV0.005 32–39 HV0.001 236–287 HV0.005

is ≤0.02 wt% C, ≤0.25 wt% Mn, 0.020 wt% P, 0.02 wt% S, and ≤0.30 wt% Ti (EN 10130). The hardness of this sheet material is typically 70–80 HV, measured at the cross-section of the sheet material. For outer parts, this material is zinc coated to ensure corrosion protection. Zinc layers are deposited by electro galvanizing (EG), hot dip galvanizing (GI), or hot dip galvanizing with additional heat treatment of the surface (galvanealled, GA). The resulting layers range in thickness from 5 to 20 µm and in hardness from about 60 to 350 HV. Furthermore, pretextured rolls are used during the final cold rolling steps in the steel plant to create specific roughness textures such as electro discharge texturing (EDT) and electron beam texturing (EBT) at the sheet surface. Characteristic values for 2D roughness, hardness, and layer thickness are given in Table 28.6. Scanning electron microscope (SEM) images and 3D roughness measurements clearly show the surface quality and texture, see respectively, Figure 28.8 and Figure 28.9 for DC 06 GI EDT. The increasing demand for light and strong structures has led to the development of new types of carbon steel sheet materials, with high(er) strength, allowing the application of thinner cross-sections in mechanical designs. Typical yield strength of these materials ranges from 500 to 700 MPa, see for example the quality HSLA in Table 28.5. This development will further broaden the spectrum of varieties in sheet surfaces. Stainless steel is another important group of sheet materials especially for domestic and household appliances such as kitchen sinks or dish washers. The chemical composition that is used strongly depends on the application but consists typically of ferritic grades like AISI 430, or austenitic grades like AISI 304. The main difference between these grades is the alloying element Ni, see Table 28.7. Appearance is crucial in marketing of household equipment. Hence, several finishes are developed by the steel manufacturers. Commonly used finishes include bright annealed (denoted as BA), annealed pickled and skin-passed (denoted as 2B), and electro discharge textured (EDT). Typical 2D roughness data measured parallel to

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28-12

Handbook of Lubrication and Tribology

FIGURE 28.8

3D roughness measurements of galvanized steel quality GI EDT.

FIGURE 28.9

SEM image of galvanized steel quality GI EDT. TABLE 28.7 Nominal Composition of Frequently Used Stainless Steel Grades in wt% Type

AISI

C

Si

Mn

Cr

Ni

Ferritic Austenitic

430 304

0.05 0.04

0.35 0.5

0.4 1.1

16.5 18.2

— 8.7

the sheet’s rolling direction is given in Table 28.8, together with an indicative value for the hardness of these materials. Sheet metal forming of stainless steel is very sensitive to galling. This is partly caused by the relatively high hardness of the material and partly because of the poor thermal conductivity of the material. Both aspects result in high local temperature rises due to frictional heating, see Section 28.4.2. Furthermore, stainless steel can have a high surface roughness, which is strongly related to wear of forming tools, especially when those tools are made of rather soft aluminum bronze, see Section 28.4.1. As the selection of the sheet material is based on the application of the formed product, it is extremely difficult to change the selected choice based on tribological considerations. As such,

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Friction and Wear in Lubricated SMF Processes

28-13

TABLE 28.8

Characteristic Roughness (µm) and Hardness Values (HV)

Grade

Finish

Ra

Rq

Ry

Rz

Rt

Rp

HV1

AISI 304 AISI 430 AISI 430

2B BA EDT

0.28 0.04 2.33

0.39 0.05 3.01

2.99 0.44 17.56

2.46 0.31 13.65

3.07 0.45 18.42

0.61 0.15 5.25

172 157 165

sheet material usually represents a fixed boundary condition in tribological optimizations of SMF processes.

28.2.4 Tool Materials Current fast changes in product design have created a need for materials designed specifically for low or medium volume series. These tool materials are usually referred to as low-cost materials or soft tool materials and are easy to process into complex shapes, saving both developing time and costs of tooling for press operations. Typical examples that have strong links with rapid prototyping techniques are epoxy resins and aluminum. Furthermore, zinc alloys (e.g., ZnAl4 Cu3 ) and resin-impregnated laminated and condensed wood can be considered. The latter material is used successfully in low-volume aerospace applications. For more demanding loading conditions, typically cast iron–carbon alloys are selected. Gray cast iron GG 25 can still be considered a low-cost material. High-volume applications that exist in mass production of large automotive panels require at least ductile cast iron grade GGG 60, but preferably GGG 70L. Mechanical strength can also be created by alloying gray cast iron with chromium, molybdenum or vanadium, see Table 28.9. For smaller tools, one could also select cast steel instead of cast iron. Cast steel has a maximum carbon content of 0.45% and small amounts of alloying elements such as manganese, molybdenum, and chromium. Frequently selected grades include 16 MnCr5 and 42CrMo4. A further increase at scale of loading conditions requires the use of wrought iron carbon alloys, usually referred to as tool steel. These steels combine the advantages of “easy” machining in the soft annealed state with high wear resistance in the hardened and tempered state. Conventional tool steel, for example, X155CrMoV 12 1/WN 1.2379 in Table 28.9, consists of 1.5 wt% C and has chromium and vanadium as main alloying elements, respectively, about 12 and 1 wt%. By changing the carbide type, size, and amount in tool steels it is possible to optimize the balance between ductility and wear resistance. Carbide volumes above 25% require a manufacturing process that is based on powder metallurgy. Typical examples of these powder metallurgical steels are high tungsten and vanadium alloyed tool steels WN 1.3344 and Vanadis 10 [10]. Powder metallurgy is also used to produce a class of cemented carbides, which is frequently denoted by the industry as hard metals, consisting of tungsten carbide bonded together by cobalt. The grain size of the tungsten carbides and the cobalt–carbide content again determines the balance between ductility, needed for impact loading, and hardness, needed for wear resistance. Full ceramics, applied for decades in conventional tribological systems, where lubrication is restricted or where extreme wear resistance is required, can also be applied in SMF, for example, as die inserts. Especially Si3 N4 [11] and SiC [12] are promising materials for highly demanding forming applications, although the poor impact strength of the materials limits its application. The range of possibilities is further enlarged by the advanced state-of-the-art coating technology. Thin hard layers can be applied on sheet and tool surfaces to meet specific demands regarding wear and corrosion resistance in a cost-effective way. A comprehensive summary of general possibilities is given by Bushan and Gupta [13], an overview of results for metal forming applications is given in Reference 9. Proven technology, extremely important in the metal forming industry, consists of heat treatment by flame or induction hardening, nitriding, and hard chromium plating. Alternative “new” technologies include single layers created by physical or chemical vapor deposition (e.g., CrN or TiN) or advanced multilayer systems (e.g., laser hardening with additional multilayer based on carbon). Chemical diffusion technologies as chromizing and vanadizing are also able to create hard, wear resistant tool surfaces.

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28-14

Handbook of Lubrication and Tribology

TABLE 28.9

Tool Materials

Examples Soft tool materials

Cast iron

Aluminum bronze

Cast steel Tool steel

Cemented carbides Ceramics

Epoxy resin Laminated wood ZnAl4Cu3 GG 25 GGG 60 GG 25 CrMoV GGG 70L CuAlFe, for example, AMPCO 25 16MnCr5 42CrMo4 WN 1.2363 WN 1.2379 WN 1.3344 Vanadis 10 94% WC–6% Co 80%WC–20% Co SiC Si3 N4

Indicative surface hardness in service 20–45 HB

Application area Low volume forming, prototype series

150–300 HB

Large tools, for example, in automotive applications

160–360 HB

Tools for deep drawing, air bending, etc. with low–medium loading conditions. Tools for stainless steel forming Tools for deep drawing, air bending, etc. with demanding loading conditions Tools for deep drawing, air bending, etc. with demanding loading conditions

600–700 HV 700–840 HV

1300–1900 HV 1500–2500 HV

Roll forming tools, tools for very demanding conditions Inserts, dry forming tools

Source: Adapted from: Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.

28.2.5 Lubrication in Sheet Metal Forming Forming lubricants are widely used in metal forming processes in order to control friction, enhance cooling, and prevent wear. The type of lubricant varies and includes: • • • •

Greases Forming oils Oil in water emulsions Protective tapes

The lubricant type and the required functionality of the lubricant is selected as a function of the forming process, see Table 28.10. The use of greases or pastes in metal forming is generally limited to low-volume applications such as the forming of panels with wooden tools or to applications at elevated temperature such as warm forming of aluminum. Typically, grease or paste is applied manually to the tooling with a brush after a series of, for example, ten products, see also Table 28.11. General multipurpose greases or pastes based on synthetic esters can be selected as forming lubricant. In many cases, it is difficult to remove the grease layers from products. An intermediate stage between greases/pastes and liquid forming oils are the so-called dry films or dry waxes. These films can be sprayed electrostatically to sheets and are usually formulated to be soluble or mixable with water so that the film can be rinsed off with water after it forms. Forming oils generally consist of a mixture of lubricant-based fluids of mineral, synthetic, or vegetable origin and additives for specific demands of metal forming applications. Additives can include, for example, oxidation inhibitors, antirust additives, foam inhibitors, viscosity modifiers, and biocides to control bacterial growth. A variety of chemical compounds is added to forming oils to meet the demanding requirements with respect to friction and wear control. These additives will create thin protective boundary layers on the sheet and tool material. The mechanism of boundary layer formation [14] can be used to classify the resulting layers, see Table 28.12. The viscosity of forming oils can be altered by viscosity

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Friction and Wear in Lubricated SMF Processes TABLE 28.10

28-15

Lubrication in Sheet Metal Forming Required functionality

SMF process Deep drawing Punch Die/blank holder Stretch forming Punch Die/blank holder Air bending Die Folding Punching Ironing

Control friction

Cooling

Prevent wear

Type of lubrication

−/
+

− −/


−/
+

Emulsion, oil, grease; in case of aluminum and stainless steel also tape

+ +

− −/


−/
+

Emulsion, oil; in case of aluminum and stainless steel also tape

−/
−/
−/
−/


− − + +



+ +

Thin oils; in case of aluminum and stainless steel also tape Emulsion, oil Emulsion

Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996. +: important,
: less important, −: not important.

TABLE 28.11 Application Methods for Forming Lubricants Water mixable

Not mixable with water

Thin emulsion

Viscous emulsion

Soap solution

Oil 70 cSt

Oil 70–350 cSt

Grease

Paste

Tape

Manual Brush/rolling Spraying Immersing Sticking

+ + + −

+

−

+
+ −

+ +
−

+ + − −

+ − − −

+ − − −

− − − +

Mechanical Rolling Dripping Spraying Immersing Adhesive

+ + + + −

+


−

+


−

+
+ + −

+ − + − −


− − − −


− − − −

− − − − +

Lubricant application method

Source: Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.

modifiers such as polymers and copolymers of olefins, methacrylates, dienes, or alkylated styrenes. The later polymers expand with increasing temperature to counteract oil thinning [15]. Heavy duty forming oils typically have a viscosity of about 300 cSt at 40◦ C. Mild forming operations of, for example, cold rolled steel can be done with forming oils of about 50 cSt or less. Some forming oils contain emulsifiers, sometimes in the form of self-emulsifying esters. Such oils can either be used directly on the sheet for severe forming applications or can be applied in the form of a oil-in-water emulsion in case of mild operations or in case cooling is required during the forming action. Finally, forming oils can be designed as vanishing oils. The base oil/solvent evaporates to the environment at a rate that depends upon the temperature, air circulation, and humidity, leaving a thin lubricant film that mainly consists of boundary layer forming additives. Vanishing oils have the advantage of reduced cleaning costs after forming. Yet, extra attention should be paid to ventilation and the risk of open flames or other sources of ignition. The boundary lubrication action of emulsions and vanishing oils is based on the same additive classes as listed in Table 28.12. Protective tapes can be applied to metal sheets or coils in order to prevent scratching during transport and storage. This kind of packaging is frequently used for aluminum and stainless steel sheet materials.

© 2006 by Taylor & Francis Group, LLC

28-16 TABLE 28.12

Handbook of Lubrication and Tribology Classification of Lubricant Boundary Layers

Boundary layer formation mechanism

Typical chemical compounds

Comments

Physical adsorption

Long chained alcohols and fatty acids

Chemical adsorption

Fatty acids, long chained fatty amides, and esters

Chemical reaction

Complex compounds based on P, S, B, and Cl

Blank feeder

Cleaning

Boundary layers consist of clustered, long chained hydrocarbons with a polar “head.” The polar group adheres to the contacting surface, by physical adsorption, forming high-viscosity hydrocarbon layers. This group of layers is meant to reduce friction and wear under mild loading conditions Boundary layers are formed by a combination of physical adsorption with chemical reaction with the surfaces, to form a metal soap. This group of layers is more resistant to increased contact temperatures, and therefore used for wear and friction reduction at moderate loading conditions Additives react with the surface to metal salts, with high temperature stability. These layers are suited for wear protection at severe loading conditions where extreme pressure (EP — additives), causes high contact temperatures. Some complex compounds are suspected to harm the environment and — without protective measures — human health

Forming lubricant

To press

FIGURE 28.10 Schematic drawing of blank cleaning and lubrication setup. Adapted from Bolt, P.J. et al., Materialen — vormgeven van dunne metaalplaat, Vereniging FME–CWM, Zoetemeer, The Netherlands, 1996.

In some cases, the protective tape is also used to prevent scratching during production and assembly of parts. The operation to remove the plastic from the sheet surface is usually performed manually, although mechanical solutions are available for specific applications. Tapes are available in a thickness range of 0.03–0.13 mm, and consist of (coextruded) polyethylene or polyvinyl chloride with an acrylic adhesive. Lubricant selection for sheet metal forming processes is a complex matter, since it involves many contradictory demands. Lubricants should, for example, adhere firmly to the sheet surface not only to maintain constant friction and constant wear prevention during the forming action but they should also be easily removable from the surface later on to avoid compatibility problems with subsequent welding or coating steps. Compatibility problems can also rise from the use of mill-applied preservation agents to protect the steel from corrosion during transport of the coils and during storage at the press shop. If such a compatibility problem exists it is necessary to use a setup as depicted in Figure 28.10. The forming lubricant is now applied after removing the preservation agent and before the actual forming operation. Such a cleaning step can be avoided in case no compatibility problems exist with the preservation agent. In that case, spot lubrication or cleaning by using washing oil is an option. Still, cleaning of the formed product is

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Friction and Wear in Lubricated SMF Processes

28-17

necessary, before a decorative coating can be applied. Cleaning becomes more and more important since many extreme pressure (EP) doped forming oils require the use of environmentally unfriendly solvents to remove boundary layers from the formed product. As such, much effort is put in dry forming or water rinseable forming lubricants.

28.3 Friction 28.3.1 Lubrication Modes Friction in lubricated systems is traditionally explained by referring to the Stribeck curve [16] and the lubrication modes related to this curve. The different lubrication modes are, respectively: • (Elasto)hydrodynamic lubrication ([E]HL). No physical contact between the asperities of the interacting surfaces occurs; the velocity difference between the surfaces is accommodated by shear in the lubricant film, which results in relative low-friction forces and, consequently, a low coefficient of friction. Typical value for f is of the order 0.01. • Boundary lubrication (BL). In this mode there is a physical contact between the asperities of the interacting surfaces. The entire load is carried by the interacting asperities. Shearing of boundary layers accommodates the velocity difference. Typically, f is in the range of 0.1 < f < 0.3. • Mixed lubrication (ML). This mode represents the intermediate regime between BL and (E)HL where a part of the load is carried by hydrodynamic action and the remaining part of the load is carried by the interacting asperities. The coefficient of friction ranges from (E)HL to BL values.

Coefficient of friction, f [-]

0.14

Boundary lubrication

Mixed lubrication

Full film lubrication

10

0.12

8

f 0.10

6

0.08 0.06

4

0.04 0.02 0.00 0.001

2

h/Ra

0 0.01

0.1 Velocity [m/sec]

FIGURE 28.11 Generalized Stribeck curve (friction) and separation.

© 2006 by Taylor & Francis Group, LLC

1

10

Film thickness/surface roughness = h/Ra [-]

It is shown that the Stribeck approach is also applicable to lubricated SMF-contacts [7]. As such, the coefficient of friction f , defined as the ratio of the friction force and the applied normal force, can be plotted in a diagram as a function of the velocity or, for instance, as a function of the dimensionless lubrication number L = ηv/(pm Ra ), introduced by Schipper [16], see Figure 28.11. The lubrication number consists of the main operational variables of study in lubricated contacts: the lubricant inlet viscosity (η), velocity (v), mean contact pressure (pm ), and the center line average roughness (Ra ). For SMF-systems it can be shown by using the operational conditions present in SMF that BL and ML will be the operational lubrication modes [7].

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Handbook of Lubrication and Tribology

28.3.2 Modeling Mixed Lubrication in Sheet Metal Forming A friction model for the ML mode can be constructed assuming that the total normal load FT acting on a contact is shared by the hydrodynamic action and the interacting asperities of the surfaces: FT = FC + FH

(28.13)

with FC being the load carried by the asperities and FH the load carried by the hydrodynamic component. 28.3.2.1 Load Carried by the Asperities The model of Greenwood and Williamson [17] can be used to estimate the load carried by the asperities: pa (x) =

2 nβσs 3




σs
E F3/2 β



h(x) σs

 (28.14)

with h being the separation between two surfaces, n the density of the asperities, β the average radius of the asperities, and σs the standard deviation of the height distribution of the asperities. F (h) is defined as: ∞ (s − h)3/2 φ(s)d s

F3/2 (h) =

(28.15)

h

where φ(s) is the height distribution of the asperity summits. Roughness is treated as a stochastic parameter with a known probability density. In this case, a Gaussian height distribution of the asperities has been assumed: 1 2 φ(s) = √ e−(1/2)s 2π

(28.16)

The elasticity modulus used in Equation (28.14) is defined as: 1 − υ12 1 − υ22 2 = + E
E1 E2

(28.17)

with Ei being the elasticity modulus of the surfaces and υi the Poisson’s ratio. The pressure distribution in rough concentrated contacts can be calculated on the basis of Equation (28.18), which is based on the work on deformation of rough line contacts presented in References 18 and 19: pC = [1 + (a1 n
a2 σs
a3 W a2 −a3 )a4 ]1/a4 pmax

(28.18)

The dimensionless numbers n
, σs
, and W , are given in Equations (28.19) and (28.20), pmax is the maximum Hertzian pressure. 32 ∗  ∗ nR βR π

(28.19)

σs
=

π σs 8 R∗

(28.20)

W =

FT BE
R ∗

(28.21)

n
=

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Friction and Wear in Lubricated SMF Processes

28-19

The values of the fit parameters ai in Equation (28.18) are: a1 = 0.953, a2 = 0.0337, a3 = −0.442, and a4 = −1.70, respectively. 28.3.2.2 Elasto Hydrodynamic Lubrication Based on calculations, Moes [20] formulated a function fit for the central film thickness in a lubricated line contact: Hc =



7/3

7/3 (3/7)s

HRI + HEI

 −1 −7/2 −7/2 (−2/7)s s + HRP + HEP

(28.22)

with s=

1 (7 + 8e−2(HEI /HRI ) ) 5

(28.23)

HRI , HEI , HRP , and HEP are the dimensionless asymptotes as described in the Moes-diagram (see Equation [28.24]): HRI = 3M −1 HRP = 1.287L 2/3 (28.24)

HEI = 2.621M −1/5 HEP = 1.311M −1/8 L −3/4 in which  E
R ∗ 1/2 η0 v + 
∗ 1/2 FT ER M=
∗ BE R η0 v + 
∗ −1/4 ER L = αE
η0 v + h H= ∗ R



(28.25)

The total pressure in an ML contact can be split into an EHL component and a BL component [21]. Figure 28.12 schematically shows the pressure distribution pT in an ML contact. In order to be able to use the results of the EHL calculations and the results of the dry contact or BL calculations, Equations (28.18) and (28.22) have to be adapted. For that purpose, two coefficients γ1 and γ2 are defined: γ1 =

FT FH

γ2 =

FT FC

(28.26)

It has been shown in References 21 and 22 that the results of the EHL and BL calculations can be combined in the ML regime, if E
is replaced by E
/γi , FT by FT /γi , and n by n · γi . For the EHL component this results in:   −7/2 7/3 −14/15 7/3 (3/7)s −7/2 (−2/7)s 1/S Hc = γ1s HRI + γ1 HEI + HRP + HEP

(28.27)

with −2/5

s = 15 (7 + 8e−2γ1

© 2006 by Taylor & Francis Group, LLC

(HEI /HRI )

)

(28.28)

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Handbook of Lubrication and Tribology

pT

pH

pC

–b

0

b x

FIGURE 28.12

Pressure distribution in a ML contact.

hs hs

dd Mean height of summits

FIGURE 28.13

Mean height of surface

The contact between a rough and a flat smooth surface is drawn schematically.

and for the BL component in: pC 1 a2 a3 = [1 + (a1 n
σs
W a2 −a3 γ2a2 )a4 ]1/a4 pmax γ2

(28.29)

As the load increases or the surface becomes smoother, pC approaches the maximum Hertzian pressure corresponding to the fraction of load carried by the asperities. From all the equations given above so far, the fractions of load of the BL component and the EHL component can be calculated. 28.3.2.3 Extension for Rough Surfaces The film thickness that is calculated with Equation (28.22) is the central film thickness of a smooth contact and is therefore, for highly loaded contacts, a measure of the volume of fluid that passes through the contact. For rough surfaces, like in SMF, a comparable measure could be found. In Figure 28.13, the contact between a rough and a flat smooth surface is drawn. In this figure, two distances between planes have been defined, that is, hs being the distance between the smooth surface and the mean plane of the summits and hs
the distance between the smooth surface and the mean plane (center line) of the rough surface. The distance between these two mean planes is given by dd : hs = hs
− dd

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(28.30)

Friction and Wear in Lubricated SMF Processes

28-21

For the extension, use is made of the definition of the film thickness given by Johnson et al. [21] as the average fluid volume between the two rough surfaces divided by the area:


hs h=

(hs
− z)φ(z)dz

(28.31)

−∞

where φ(z) is the distribution of surface heights. For situations with a large film thickness, the average film thickness is the same in both models. This interpretation of h improves the prediction of the mixed lubrication model. Basically, for a given load partition, γ1 and γ2 , the hydrodynamic load yields the central film thickness, Equation (28.22), from which the value of h = hs giving the asperity pressure to the central pressure given directly by the asperity pressure, Equation (28.14), can be found. Equating this pressure to the central pressure given directly by the asperity load, Equation (28.29) is the criterion used to provide the self-consistent load partitioning of the problem [23].

28.3.3 Calculation of the Stribeck Curve The total friction force Ff is the sum of the friction force between the interacting asperities and the shear forces of the hydrodynamic component:

Ff =

N 

 τCi dACi +

i=1 A Ci

τH dAH

(28.32)

AH

with N the number of asperities in contact ACi the area of contact of a single asperity, i; τCi the shear stress at the asperity contact i; AH the contact area of the hydrodynamic component; and τH the shear stress of the hydrodynamic component. The coefficient of friction fCi of a single asperity can be written as: fCi =

τCi pCi

(28.33)

with pCi the normal pressure of a single asperity. Briscoe et al. [24] showed that the ratio of the shear strength and the local contact pressure is nearly constant. Since the coefficient of friction is by approximation constant for all asperity contacts, the first term in Equation (28.32) can also be written as: N 

τCi dACi = fC FC

(28.34)

i=1 A Ci

where FC is the total load carried by all asperities. The value of fC is determined from experiments. For the shear force in the film different models can be used, for example, the isothermal Eyring model: 

ηγ˙ τH (γ˙ ) = τ0 · arc sinh τ0

 (28.35)

where η is calculated according to the Roelands equation, the pressure used to calculate the viscosity is the average pressure of the hydrodynamic component, that is, the total load carried by the hydrodynamic component divided by the total hydrodynamic contact area. γ˙ is the shear rate (v dif /h) and τ0 is the Eyring shear stress. v dif is the velocity difference between the two surfaces.

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28-22

Handbook of Lubrication and Tribology

Input (FT, v, h, ...)

Assume FC Eq. 3.1 / Eq. 3.14

Eq. 3.14

FH g1

g2

Eq. 3.15

Eq. 3.17

HC

PC Eq. 3.19

hs’ Eq. 3.18 hs Eq. 3.2 Pa(0) Equal? No

Yes, Eq. 3.25 f

FIGURE 28.14

Schematic representation of Stribeck curve calculation.

The coefficient of friction can thus be written as: fC FC + Ff = f = FT



AH τH (γ˙ )dAH

FT

(28.36)

Combining this result with Equation (28.35) results in the coefficient of friction for Stribeck curves: f =

fC FC + τ0 AH arc sinh(ηv dif /hτ0 ) FT

(28.37)

The Stribeck curve for SMF-contacts can now be calculated as demonstrated in Figure 28.14 by varying the sliding velocity. In Table 28.13 the experimental conditions of measurements performed by ter Haar [7] are given. In this table the value for Ra and dd are taken from actual measurements. The experiments were performed to simulate friction in sheet metal forming and to establish the Stribeck curve for sheet metal forming. The results of the measurements of ter Haar are plotted in Figure 28.15, together with calculations with the present model. Two different lubricants are used for the experiments. The presented calculations are done with the thicker oil of the two, calculations with the thinner oil is only slightly different. The friction in the (E)HL regime is slightly lower for Lub1.

28.4 Wear Tool wear and galling are the two main wear types that limit tool life in sheet metal forming applications. Based on the classification scheme presented in [25] both types can be defined as special cases of the

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Friction and Wear in Lubricated SMF Processes

28-23

TABLE 28.13 Experimental Conditions for UCS1-Material Property

Value 7.7 109 m−2 4.21 µm 2.19 µm 30 mm 231 GPa 50 mm 1.125 µm 1.85 µm 0.6 Pa sec (Lub1) 1.2 Pa sec (Lub2) 3.3 10−8 Pa−1 0.68 196.2 MPa 2.5 MPa 0.13 0.25% 350 N 72.7 MPa

n β σs B E
R dd Ra η0 η0 α z pr τ0 fC Sep FT pav

Source: Adapted from ter Haar, R., Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente, 1996.

0.16

Coefficient of friction, f [-]

0.14 0.12 0.10 0.08 0.06 0.04 0.02 0.00 10–5

Lub 1 Lub 2 Present model

10–4

10–3

10–2

Lubrication number, L [-]

FIGURE 28.15 Measured and calculated Stribeck curve. Adapted from Gelinck, E.R.M., Mixed lubrication of line contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands, 1999.

sliding wear process (see also Reference 9). The velocity difference between the sheet and the tool during a forming action should preferably be accommodated by shearing of an interfacial layer, for example, a lubricant film. However, in some cases either the tool surface or the sheet surface is damaged due to the sliding action. The first case causes volumetric wear of the tool. The dominant wear mechanism in the sense of Reference 25 is abrasive wear. Reshaping of the tool surface is necessary in order to avoid products with unacceptable dimensions. The second case is characterized by severe scratching of the sheet

© 2006 by Taylor & Francis Group, LLC

28-24

Handbook of Lubrication and Tribology 15000 Lifetime [amount of products]

R = 0.7 mm

12500

Lifetime criterion: radius increase of 1 mm

10000 7500

GI

EZ

GA

5000 2500 0 Epoxy resin

Laminated wood

ZnAl4Cu3

GG-30

FIGURE 28.16 The lifetime of a forming tool as a function of the tool material for different zinc coated steel sheets. The triangular shaped area indicated by the arrow is sensitive to wear. The amount of products needed for a 1 mm increase of the top radius is used as lifetime criterion. Adapted from van der Heide, E. et al., Journal of Materials Processing Technology, 141/2, 197–201, 2003.

surface and is actually a form of volumetric wear of the sheet surface. This wear type is generally referred to as galling, and is related to pick-up of sheet material at the tool’s surface, a combination of abrasive and adhesive wear. Polishing of the tool surface is necessary in order to avoid an unacceptable surface condition of the product and extreme friction forces during forming.

28.4.1 Volumetric Wear of Forming Tools Volumetric wear of a forming tool is first visible at spots with high local pressure. Take for example, the tool given in Figure 28.16, which produces a box shaped product of 100 × 100 mm with a depth of 20 mm. The part sensitive to volumetric wear is indicated by the arrow. Results presented in Reference 26 show large differences in lifetime for this tool geometry as a function of the applied (soft) tool material and as a function of the processed sheet material. Clearly this points the system’s dependence of wear. A common approach in wear assessment is based on establishing a relation between volumetric wear and the operational conditions. It is shown, for example, by Reference 27 that an Archard type of wear equation [28], here presented in the nondimensional form Kw [29] by Equation (28.38) and in terms of the specific wear rate k by Equation (28.39), could be used for this purpose. Kw =

k=

V s · FN

V · Hsoft s · FN

mm3 (Nm)−1

(28.38)

(28.39)

in which Kw represents the coefficient of wear, V the wear volume, s the sliding distance, FN the applied normal force, and Hsoft the hardness of the softest contact partner. Achard’s wear equation has proven its value in comparative wear testing. But even more important, it can be used in the design environment based on the finite element method that is used extensively in sheet metal forming applications. One could, for example, build a routine that adjusts the geometry of the contacting tool surface based on the operational conditions that were applied on the surface during the forming step prior to the one of interest that is: the length of the sheet surface in contact with the element and the average normal force applied on the element during the forming action. The remaining task for a tribologist is to produce reliable data for the specific wear rate [26,30]. The simplest first guess is related to the hardness of the tool material: wear resistance increases with increasing hardness of the tool material. Equation (28.38) suggests the same: volume loss of the forming tool is inversely proportional to the hardness of the softest material. Abrasive wear tests like the rubber

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Friction and Wear in Lubricated SMF Processes

28-25

TABLE 28.14 Specific Wear Rate of Conventional Tool Materials Measured with Rubber Wheel Abrasion Test ASTM G65 at Fn = 100 N Using Quartz Sand as Abrasive (900 ± 50 HV, Grain Size 0.2–0.5 mm, Rounded Shape). WN 1.2379 is Hardened and Tempered to 60 HRc GGG 40 k [10−6 mm3 /Nm]

GGG 60

GGG 70L

WN 1.2379

1030

859

46

1357

Source: Adapted from van der Heide, E. et al., in Proc. NORDTRIB 04 Conference, Tønder, K. et al., Eds, Tromsø, Norway, 2004, pp. 355–364.

Volumetric wear [mm3]

0.50 0.40 0.30 0.20 0.10 0.00 0

20

40

60

80

100

120

140

Distance [m]

FIGURE 28.17 Rubber wheel abrasion test results, for hard chromium plated GGG 60. The specific wear rate increases significantly after the 50 µm Cr layer has worn off. TABLE 28.15

Slider-On-Sheet Test Results. FN = 100 N

Tool material Aluminum bronze Aluminum bronze GG-25 Aluminum bronze ZnAl4 Cu3 ZnAl4 Cu3 Epoxy resin

Sheet material

Lubricant

k-value 10−6 mm3 /Nm

AISI 430 BA AISI 430 BA DC 04 GA AISI 304 2B DC 04 GI DC 04 GA DC 04 GA

42 K 140-EL N 6130 140-EL N 6130 N 6130 N 6130

0.020 0.051 0.72 0.95 2.41 89.5 1450

wheel abrasion test or the Taber abrader test, confirm this, see for example Table 28.14, in which the specific wear rate is given for three commonly used cast irons and hardened tool steel WN 1.2379 (60 HRc). A further increase in hardness can be reached by the application of thin hard coatings like hard chromium plating or physical vapor deposition of CrN. The same rubber wheel test now shows a substantial decrease in worn volumes, until the layer has worn off, see Figure 28.17 for hard chromium plated GGG 60. Although the specific wear rates measured with the rubber wheel test or similar conventional wear tests can be used as a comparative measure of performance, it is clear that the calculated k-values cannot be used as input to FEM sheet metal forming applications. As such, dedicated test methods have been designed that take into account the tribological system that is used in lubricated sheet metal forming processes. Results with, for example, a slider-on-sheet configuration indicate the importance of measuring the wear response of materials in relation to the tribological system in which it is used. The specific wear rate can easily vary by a factor of 100 from one system to the other. Especially, the application of rough stainless steel can increase the measured wear rate substantially. For lubricated smooth stainless steel, specific wear rates in the order of 10−8 mm3 Nm−1 are found. For rough qualities this could easily be 10−6 mm3 Nm−1

© 2006 by Taylor & Francis Group, LLC

28-26

Handbook of Lubrication and Tribology TABLE 28.16 Process Characteristics Description

Value

Dimensions of the blank Dimensions of the die Die inner diameter Draw radius Punch force Blankholder force

Ø 440 × 2.1 mm Ø 440 × 80 mm Ø 257 mm 10 mm 800 kN 300 kN

Scratch depth [mm]

25 20 15 10 5 0 0

50

100

150

200

250

Scratch width [mm]

FIGURE 28.18

Depth and width of unacceptable scratches on an axisymmetric deep draw product.

depending on the lubricant used [30]. The same holds for the introduction of hard zinc layers like the galvanealled (GA) quality, see Table 28.15.

28.4.2 Galling The importance of control of galling can be illustrated by the production of 18 l expansion tanks for central heating systems. The upper and lower halves of the tank are made from 2.1 mm thick, cold rolled steel DC 03, by axisymetric deep drawing (see Table 28.16). A forming lubricant is applied on both sides of the blanks. The side in contact with the drawing die uses 24 kg of lubricant per 10,000 products. Cast steel 42CrMo4 , uncoated and not hardened, is used as tool material for the drawing die. During production, scratches appear on the products surface. The severity of scratching increases with the amount of products and reach an unacceptable level after 3,400 products, see Figure 28.18. Galling, in the context of metal forming, is associated with the tendency for lubricant film breakdown, resulting in pick-up of sheet material by the tool surface and subsequent scoring (severe scratching) of the work piece surface [31]. Scoring or severe scratching may, by definition, be due to local solid-phase welding or to abrasion. Galling mechanisms in SMF operations can be divided into three phases [32,33]: 1. Initiation 2. Lump growth 3. Severe scratching or seizure The initiation of material transfer occurs at tool surface defects like grinding marks or carbides. This can be understood by taking into account the contact situation given in Figure 28.19. As stated in Section 28.2.2, a hard and smooth tool surface interacts with a relatively soft and rough sheet material. The sheet material will plastically deform by applying normal force to the tool–work piece interface. Consequently, the sheet’s surface roughness will change and form roughness plateaus. Tool summits or surface defects, see Figure 28.20, now become important because they will plough through the plastically deformed

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Friction and Wear in Lubricated SMF Processes

28-27

Tool surface

hs Mean plane surface heights S Mean plane summit heights

FIGURE 28.19 Multi-asperity contact between a rigid (tool) surface and a plastically deforming (sheet) surface. Adapted from van der Heide, E. and Schipper, D.J., ASME Journal of Tribology, 126, 2, 275–280, 2004.

FIGURE 28.20 SEM image of roughness summits/surface defects on a tool steel surface. Adapted from Heikillä, I. et al., Proc. Innovations in Metal Foming, Brescia, Italy, 2004.

roughness plateaus at the sheet surface as a result of the sliding action, see Figure 28.19. The sliding action will, depending on the interfacial shear strength, the size of the summit, and the summits attack angle result in ploughing, cutting, or wedge formation [34]. Tool surface defects that operate in the wedge formation mode will initiate galling in unlubricated sheet metal forming contacts [35]. The interaction of tool summits and the work piece surface generates heat, which in turn results in a local surface temperature rise. The relation between frictional heating and surface temperature rise at the asperity level can be described by an extension of the flash temperature model, developed for ideally smooth surfaces by Bos [36]. From calculations [37,38] it follows that the flash temperature Tf at the summit level, strongly depends on the hardness of the sheet material, the thermal properties of the sheet material (conductivity and diffusivity), the operational conditions (load and velocity), and the thermal conductivity of the tool material. Galling initiation in lubricated sheet metal forming processes will occur in case the protective lubricant boundary layer, present at the interface of the tool summit–work piece contact, fails. The latter occurs, depending on the concentration of boundary layer forming compounds and depending on the boundary layer formation mechanisms [39,40], at a certain critical temperature, Tcr [41]. The model for frictional heating at the asperity level, can be applied to SMF tool surfaces, by calculating the relevant input parameters of the model for each individual tool summit. Introduction of

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the constancy of scoring temperature and the wedge formation mode as conditions for material transfer yields a tool for prediction of galling initiation. Based on model simulations, one can obtain two general strategies for avoiding the occurrence of galling [37]: 1. Exclude the occurrence of wedge formation, the necessary condition for material transfer, by: • The application of surfaces, for example, coatings, that introduce a low interfacial shear strength • Removal of possibly wedge forming summits, meaning not only the highest summits that operate in the cutting regime, but also the summits of intermediate height (mirror polish) 2. Exclude high flash temperatures, the necessary condition for lubricant failure, by: • The application of lubricants that resist high temperature • The application of tool materials/tool coatings with high thermal conductivity, which will decrease the maximum surface temperature rise to values below Tcr • The application of soft sheet materials, since a decrease in hardness causes a significant decrease in Tf , possibly below Tcr The first and fifth options are not as straight forward as the other options, because they require to change the sheet material, which in turn could influence the diffusivity of the material or the available area for heat conduction. Further, changes of the sheet material are generally impossible to implement, due to the specific demands of the application. The combination of a smooth tool surface with high thermal conductivity, however, is shown to be successful: galling initiation is avoided for a typical sliding contact with stainless steel sheet [42,9]. The second phase in galling processes, stresses the accumulative nature of galling: lumps on the tool surface grow as a function of the amount of products formed. Lump growth is controlled by the probability of wedge formation in the contact between a tool surface asperity and the flattened sheet asperities. A SEM image of a lump on tool steel after sliding contact with stainless steel sheet material is given in Figure 28.21. The mechanisms that control the growth rate can be simulated numerically for unlubricated sheet metal forming processes [35]. Lump growth continues to a certain point where the lumps reach a critical size and shape, which results in severe damage to the sheet surface. This third and final phase, more stochastic in nature than the previous phase, is characterized by severe scratching of the sheet material and possibly ends with seizure. The effect of lump growth on the severity of scratching of the sheet surface and its relation with friction is

WN 1.2379

AISI 304

FIGURE 28.21 Lump growth on tool steel WN 1.2379 after sliding contact with stainless steel sheet material AISI 304 2B. Adapted from Lovato, G. et al., High volume forming of stainless steel with easy to clean lubricants, ECSC — steel report, contract number 7210-PR-307, 2003.

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Scratch depth (mm)

Coefficient of friction [-]

Scuffing

lump growth

Initiation

Sliding length [m]

FIGURE 28.22 The effect of lump growth on the severity of scratching and friction. Adapted from van der Heide, E., Huis in ‘t Veld, A.J., and Schipper, D.J., Wear, 251–12, 973–979, 2001. TABLE 28.17 Results of Hat Drawing Tests with Coated Tools and Different Lubricants. The Maximum Test Run was 20,000 Hats. Sheet Material AISI 304 (EDT-) Skin Passed

WN 1.2379 WN 1.2379 WN 1.2379

Treatment

Lub. A 71 cSt

Lub. B 190 cSt

Lub. C 160 cSt

Lub. D 80 cSt

Lub. E 36 cSt

Hardened CVD TiN PVD TiCN

20000 20000 20000

2400 4300 20000

2000 20000 20000

500 19000 2000

500 2000 200

Lub. A, Cl containing reference lubricant. Lub. B–D, alternative lubricants (no Cl). Lub. E, preservation oil for DC type of sheet. Source: Adapted from Jordan, F. and Heidbuchel, P., IDDRG Working Group Meetings, IDDRG Birmingham, UK, June 1999.

given in Figure 28.22, which is based on slider-on-sheet test results [43]. It can be seen from Figure 28.22, that lump growth on the die results in increased depth of scratches on the sheet. A similar relation between scratch depth and lump growth is found, for example, by Murakawa [44] for a deep drawing application of aluminum. The selected forming lubricant largely influences the growth rate of lumps on the tool. It has been shown in forming trials, see for example, Table 28.17, with sheet materials sensitive to galling, that effective boundary lubrication delays or even prevents severe scratching.

28.5 Simulation of Tribological Contact Situations in Sheet Metal Forming 28.5.1 General Framework of Demands for Tribo Tests The effect of changes in SMF-contacts, like the application of surface treatment, coating technology, or enhanced lubricant chemistry, on tool life and product quality, should preferably be measured by actually performing tests at an industrial scale. Practical reasons like the unavailability of a trail press or lack of time and money to perform extensive trails, has created the need for tribological screening tests. In general, tribological testing is performed, assuming that industrial applications can be reduced to contact situations with input and output as illustrated by the system approach in Section 28.2. The extracted system is then simulated at a laboratory scale using the same system components and input. Since structure and input

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are similar, it is assumed that the output of the experiment, that is, friction and wear, correlates with the tribological performance in industrial practice [9]. Wear and friction tests related to SMF differ from conventional wear tests because of the specific system components involved and the required type of motion. Actually, the following requirements should be met [43]: 1. Since sheet-tool interaction has to be simulated, it is essential that a well-defined, reproducible contact between sheet and tool material is maintained during the test. 2. It is essential that the sheet material in contact is always fresh. Naturally, the general requirements for a tribotest should be obeyed, that is 3. The operational variables should be applied as in industrial practice. 4. Wear and friction should be measured directly. 5. Friction should be measured independently of the normal force. Because the method should also be economically feasible, final requirements are: 6. The test pieces used should be “easy” to make. 7. The test method should allow for sliding distances in the kilometer range, in case wear is the object of study.

28.5.2 Conventional Tribo Tests and Sheet Metal Forming The first two of the requirements of Section 28.5.1 imply that modified conventional tribo testers, like rotational and reciprocating testing devices, could only generate data linked to SMF for one turn or stroke. This type of testing is used often to measure the effect of sheet roughness and lubricant viscosity on friction as a function of sliding velocity and normal pressure, see for example, Reference 45. This kind of equipment is useful for generating frictional data for input of FEM sheet metal forming simulations, if the test pieces and the operational variables resemble the industrial application of interest (requirements 1 and 3). Wear measurements cannot be done because of the limited sliding distance. Yet, much tribological work related to lubricants is still done at conventional tribo tester like the Shell four ball machine, the Falex pin & vee block test machine or the Timken block-on-ring tester. Typically, one changes the standard test piece material to the materials used for the practical application, such as an aluminum ring vs. a tool steel block, and simply measures the extreme pressure characteristics of the lubricant of interest [2]. This kind of testing is able to rank two lubricants, one with and one without extreme pressure additives like chlorinated paraffin, correctly. Rankings found for a set of commercially available forming oils however, will have very limited value for sheet metal forming applications.

28.5.3 Simulation of Friction and Wear in Sheet Metal Forming Frequently used tribo tests for SMF-conditions make use of strip material drawn between two clamped cylinders, flat dies (flat die test), or over a simple radius (strip stretching test), see Figure 28.23. A strip of sheet material is pulled over a radius, during the strip stretching test, applying a constant pulling force and pulling velocity at point 1. The other end of the strip, point 2, is fixed. This test is also carried out allowing for linear displacement of both ends of the strip (radial strip drawing test). The back-pull force can be used to increase the normal force acting on the radius [46]. The flat die test geometry resembles the contact situation that exists in blank holder–work piece interfaces. The flat die test is typically used to generate friction data in the mixed or boundary lubrication regime, at different contact pressures. A modified test sequence is sometimes used to assess the galling characteristics of zinc coated steel sheet. A specific sequence referred to as the multi frottement test, is used by major European steel sheet manufacturers. The steel strip is oiled once for this test and pulled several times between two flat dies until galling appears. A maximum of 10 strokes of 150 mm is made in this sequence but, to increase the severity of the test, one can opt to use longer strokes or to pull a larger number of strokes.

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FBP Fp 2

Fn

Fn

1 Fp

FIGURE 28.23 Schematic drawing of the Strip Stretching Test setup (left) and the Flat Die Test setup (right). Adapted from Sniekers, R.J.J.M., Friction in deep drawing, PhD Thesis, Eindhoven Technical University, 1996.

Fp 1

2

FIGURE 28.24 Schematic drawing of the Draw Bead Test setup. Adapted from Sniekers, R.J.J.M., Friction in deep drawing, PhD Thesis, Eindhoven Technical University, 1996.

Sheet specimen Spring blades Elastic joint

Sliding tool Bellows Support friction force transducer

Rotating tool Support

Normal force transducer

Main support

FIGURE 28.25 Friction device for the cylinder on strip tribometer. Adapted from de Rooij, M.B., Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente, 1998.

Draw bead tests, can be regarded as a variant of this type of testing. Strip material is drawn through a set of three cylinders, as shown in Figure 28.24, by imposing a horizontal displacement at point 2, while suppressing vertical displacement at point 1 [46]. Tests are performed with cylinders that can rotate freely or with fixed cylinders. Again, this test is typically done to generate friction data for simulation of

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Plane surface test

Cylinder-plane test

Blank holder

Sheet metal

Die

Draw bead test

Cylindrical blank holder

Sheet metal

Die

Blank holder

Sheet metal

Draw bead

Die

Tool Working grab

Cam discs Blank holder lever

Drive train

FIGURE 28.26 PtU strip drawing test with exchangeable tool unit. Adapted from Groche, P., Filzek, J., and Nitzsche, G., in Wissenschaftliche Gesellschaft für Produktionstechnik (WGP): Annals of the German Academic Society for Production Engineering, Braunschweig, XI/1, Vol. 1, 2004, pp. 55–60.

the material flow near draw beads. In order to asses galling tendencies, one can for example apply the so-called multi-strip galling test sequence, which was developed for low carbon steel. This test method consists of pulling one oiled strip through the draw bead setup followed by the pulling of fresh nonoiled strips through the beads until galling occurs. A first indication of the galling tendency can be deduced from the number of strips that can be applied without the appearance of visual scratches on both sides of the strip. Friction can be measured direct and independent of the normal force, using a rotating tool in the cylinder on strip method shown in Figure 28.25 [7]. The sliding tool is used for friction measurements, the rotating tool for support during the test. By using a tensile tester for clamping the strip material, experiments can be conducted under conditions of controlled (plastic) deformation of the strip. A well-accepted intermediate stage between practice and laboratory testing uses a rather complex test piece geometry involving bending over a radius [47,48] or drawing a strip between two stationary test pieces, respectively simulating the sheet–tool contact at the die radius and the blank holder contact during

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y FN

v x

FIGURE 28.27

I

Schematic representation of the TNO slider-on-sheet test.

deep drawing. Recently a new PtU-strip drawing tester was designed in which a slitted coil is fed to an exchangeable tool unit [49]. Figure 28.26 shows the test set up and three possible tool units. The upper and lower tool units are equipped with piezoelectric force transducers in order to measure the normal load and friction force separately on each tool. The normal force is applied by means of a lever and a package of leaf springs. The sliding velocity can be adjusted by the rotational speed of the motor. One stroke of the intermitting test is typically 100 mm, a maximum of 15,000 strokes is performed for a standard wear test. Second, long sliding length experiments can be performed with the TNO slider-on-sheet tribo meter [43], see Figure 28.27. This tribo meter consists of a sliding contact between a ring, made of the tool material of interest and sheet material used in the application. Each track is made next to the previous track, in the same direction with sliding velocity v and under a normal load of Fn , thus assuring virgin sheet material in the contact. If the tracks are made 1 mm apart from each other, it is possible to realize 1 km sliding distance on 1 m2 sheet material. The friction force is measured with the help of a force transducer with strain gauges and a cantilever beam construction. The normal force is applied by means of pressurized air. The sliding velocity (0.001–1 m/sec), the normal force (50–1000 N) and the track length (100–2000 mm) can be adjusted within their ranges.

Acknowledgments The authors wish to thank the following people for the contributions to the work: Dr. ir. E.R.M. Gelinck at TNO for his input for Section 28.3 Friction and Dipl.-Ing. G. Nitzsche at PtU Darmstadt for his kind cooperation and supply of Figure 28.5 and Figure 28.26.

References [1] ASM Handbook, Vol. 14, Forming and Forging; ASM: Materials Park, OH, USA. [2] Schey, J.A., 1970, Metal Deformation Processes Friction and Lubrication, Marcel Dekker Inc., New York. [3] Czichos, H., 1978, Tribology: A Systems Approach to the Science and Technology of Friction, Wear and Lubrication, Tribology series, Vol. 1, Elsevier Scientific Publishing Company, Amsterdam. [4] Salomon, G., 1974, Application of systems thinking to tribology, ASLE Transactions, 17, 295–299. [5] Halling, J., 1986, The tribology of surface coatings, particularly ceramics Proceedings of the Institution of Mechanical Engineers, 200, C1, 31–40. [6] Greenwood, J.A. and Williamson, J.B.P., 1966, Contact of nominally flat surfaces, Proceedings of the Royal Society of London A 295, 300–319. [7] ter Haar, R., 1996, Friction in sheet metal forming: the influence of (local) contact conditions and deformation, PhD Thesis, University of Twente. [8] de Vin, L.J. et al., 1996, A process model for air bending, Journal Materials Processing Technology, 57/1–2, 48–54.

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[9] van der Heide, E. and Schipper, D.J., 2004, Tribology of Metal Forming, in Mechanical Tribology: Materials, Characterization, and Applications, Totten, G.E. and Liang, H., eds, Marcel Dekker Inc., New York, pp. 347–374. [10] Steels for Cold Work Tooling, 2001, Uddeholm Tooling, Hagfors, Sweden. [11] Doege, E. and Dröder, K., 1997, Einsatz von Keramik als Werkzeugwerkstoff in der Blechumformung, Bänder Bleche Rohre, 12, 16–21. [12] Kataoka, S. et al., 2004, Tribology of dry deep-drawing of various metal sheets with use of ceramics tool, Surface and Coatings Technology, 177–178, 582–590. [13] Bushan, B. and Gupta, B.K., 1991, Handbook of Tribology — Materials, Coatings and Surface Treatments, McGraw-Hill, New York. [14] Bowden, F.P. and Tabor, D., 1950, The Friction and Lubrication of Solids, Oxford, Clarendon Press. [15] www.lubrizol.com. [16] Schipper, D.J., 1988, Transitions in the lubrication of concentrated contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands. [17] Greenwood, J.A. and Williamson, J.B.P., 1996, Contact of nominally flat surfaces, Philosophical Transactions of the Royal Society of London A, 295, 300–319, [18] Gelinck, E.R.M., 1999, Mixed lubrication of line contacts, PhD. Thesis, University of Twente, Enschede, The Netherlands. [19] Gelinck, E.R.M. and Schipper, D.J., 1999, Deformation of rough line contacts, ASME Journal of Tribology, 121, 3, 449–454. [20] Moes, H., 2000, Lubrication and Beyond, Lecture Notes, University of Twente, Enschede, The Netherlands. [21] Johnson, K.L., Greenwood, J.A., and Poon, S.Y., 1972, A simple theory of asperity contact in elastohydrodynamic lubrication, Wear, 19, 91–108. [22] Gelinck, E.R.M. and Schipper, D.J., 2000, Calculation of Stribeck curves for line contacts, Tribology International, 33, 175–181. [23] Gelinck, E.R.M. and Schipper, D.J., 2001, Stribeck and traction curves for highly loaded contacts, Report, University of Twente, TR-012263. [24] Briscoe, B.J., Scruton, B. and Willis, F.R., 1973, The shear strength of thin lubricant films, Philosophical Transactions of the Royal Society of London A, 333, 99–114. [25] DIN 50320, 1979, Verschleiß — Begriffe, Systemanalyse von Verschleißvorgängen, Gliederung des Verschleißgebietes, Beuth Verlag, Berlin. [26] van der Heide, E. et al., 2003, Wear of soft tool materials in sliding contact with zinc coated steel sheet, Journal of Materials Processing Technology, 141/2, 197–201. [27] Eriksen, M. and Wanheim, T., 1997, Wear optimisation in deep drawing, in Proceedings of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 128–133. [28] Archard, J.F., 1953, Contact and rubbing of flat surfaces, Journal of Applied Physics, 24, 981–988. [29] Shaw, M.C., 1977, Dimensional analysis for wear systems, Wear, 43, 263–266. [30] van der Heide, E. et al., 2004, Wear of aluminium bronze in sliding contact with lubricated stainless steel sheet material, in Proceedings NORDTRIB 04 Conference, Tønder, K. et al., eds, Tromsø, Norway, pp. 355–364. [31] Andreasen, J.L., Eriksen, M., and Bay, N., 1997, Major process parameters affecting limits of lubrication in deep drawing of stainless steel, in Proceedings of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 122–127. [32] Schedin, E. and Lehtinen, B., 1993, Galling mechanisms in lubricated systems: a study of sheet metal forming, Wear, 170, 119–130. [33] Schedin, E., 1994, Galling mechanisms in sheet forming operations, Wear, 179, 123–128. [34] Hokkirigawa, K. and Kato, K., 1988, An experimental and theoretical investigation of ploughing, cutting and wedge formation during abrasive wear, Tribology International, 21, 1, 51–57.

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[35] de Rooij, M.B., 1998, Tribological aspects of unlubricated deepdrawing processes, PhD Thesis, University of Twente. [36] Bos, J. and Moes, H., 1995, Frictional heating of tribological contacts, Journal of Tribology, 117, 171–177. [37] van der Heide, E., 2002, Lubricant failure in sheet metal forming processes, PhD Thesis, University of Twente, Enschede, The Netherlands. [38] van der Heide, E. and Schipper, D.J., 2004, On frictional heating in single summit contacts: towards failure at asperity level in lubricated systems, ASME Journal of Tribology, 126, 2, 275–280. [39] Frewing, J.J., 1943, The heat of adsorption of long-chain compounds and their effect on boundary lubrication, Proceedings of the Royal Society of London, A, 182, 270–285. [40] Spikes, H.A. and Cameron, A., 1973, Scuffing as a desorption process — an explanation of the Borsoff effect, ASLE Transactions, 17, 2, 92–96. [41] Blok, H., 1969, The postulate about the constancy of scoring temperature, in Interdisciplinary Approach to Friction and Wear, Ku, P.M., ed., Symposium Troy, New York, NASA SP-237, pp. 153– 248. [42] van der Heide, E. and Schipper, D.J., 2003, Galling initiation due to frictional heating, Wear, 254/11, 1127–1133. [43] van der Heide, E., Huis in ‘t Veld, A.J., and Schipper, D.J., 2001, The effect of lubricant selection on galling in a model wear test, Wear, 251–12, 973–979. [44] Murakawa, M., Koga, N., and Takeuchi, S., 1997, Utility of diamondlike carbon-coated dies as applied to deep drawing of aluminum sheets, in Proceeding of the 1st International Conference on Tribology in Manufacturing Processes ’97, Dohda, K., Nakamura, T., and Wilson, W.R.D., eds, Gifu, Japan, pp. 322–327. [45] Emmens, W.C., 1997, Tribology of flat contacts and it application in deep drawing, PhD Thesis, University of Twente. [46] Sniekers, R.J.J.M., 1996, Friction in deep drawing, PhD Thesis, Eindhoven Technical University. [47] Woska, R., 1982, Einfluß ausgewählter Oberflächenschichten auf das Reib- und Verschleißverhalten beim Tiefziehen. PhD Thesis, TU Darmstadt. [48] Schulz, A. et al., 1997, Deposition of TiN PVD coatings on cast steel forming tools, Surface and Coatings Technology, 94–95, 446–450. [49] Groche, P., Filzek, J., and Nitzsche, G., 2004, Local contact conditions in sheet metal forming and their simulation in laboratory test methods, in Wissenschaftliche Gesellschaft für Produktionstechnik (WGP): Annals of the German Academic Society for Production Engineering, Braunschweig, XI/1, Vol. 1, pp. 55–60.

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Maintenance

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III

29 The Degradation of Lubricants in Service Use 29.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29-3

Controlled Deterioration of Lubricants • The Effects of Deterioration • Physical Causes of Deterioration • The Effects of Lubricant Chemical Deterioration • The “Bath-Tub Curve”

29.2 Field Tests for Lubricant Deterioration . . . . . . . . . . . . . .

29-8

Direct Observation of Lubricant Condition • Field Kits for Lubricant Condition

29.3 Laboratory Tests for Lubricant Deterioration . . . . . . .

29-10

Viscosity and Viscosity Index • Trace Metals • Particulates and Ash in Lubricants • Acidity and Base Reserve • Water Content

29.4 Minor Methods of Investigating Lubricant Degradation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29-33

Density, “Gravity,” or “Specific Gravity” • Flash Point of Degraded Lubricant • Foaming of Lubricants • System Corrosion (“Rusting”) with Degraded Lubricants • Demulsibility and Interfacial Tension of Degraded Lubricants • Instrumental Analytical Techniques

29.5 Case Studies of Degraded Lubricants . . . . . . . . . . . . . . . .

29-36

A Degraded Lubricant Sample from a Heavy Duty Diesel Engine • A Degraded Grease Sample • A Degraded Lubricant Sample from a Gas-Fueled Engine • A Degraded Hydraulic Fluid • Overview of Degraded Lubricant Analyses

Malcolm F. Fox De Montfort University

Bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Reference . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29-40 29-40

29.1 Introduction The very nature of lubricant service means that lubricants deteriorate during their service use. It is normal that lubricants degrade by partial evaporation, oxidation, and contamination. The purpose of lubricant formulation for a defined application is to control the deterioration of that lubricant in a planned manner over an established period of time, work, distance, or operation. The deterioration of a lubricant can either 29-3

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be planned and controlled by various means or be uncontrolled. Modern practice is strongly directed to the former.

29.1.1 Controlled Deterioration of Lubricants The way that a lubricant is changed in service use addresses the two extremes of one of the following: • A time- or distance-defined period of lubricant replacement, such as 500 h operation, annually or 10,000 km, without regard to the actual state of the lubricant. But custom and practice show that the service interval set is sufficient to ensure that excessive wear does not occur — a precautionary principle. This approach does not require sampling and analyses or “on-board” sensors and therefore of low-cost. The issue is that the lubricant is replaced with a substantial amount of remaining “life” in it, therefore tending to be wasteful of resources. • At the other extreme, a quantitative appreciation of the state of the lubricant is done by sampling at regular intervals and monitoring various parameters to give a collective assessment of the condition of the lubricant, “condition monitoring.” The time interval of sampling should be, at most, half of the anticipated service interval. The database built up over time has value for long term and is concerned with long-term trends in lubricant parameters such as wear metal concentrations, viscosity, and particulate levels. For a full condition monitoring program, the lubricant is replaced when its condition reaches a lower bound of aggregated parameters and it is judged to be, or close to being, unsuitable for its purpose of lubricating and protecting the mechanical system. • An interim position is to sum the overall performance of the system, be it engine or machine, from its last service interval by integrating power levels used in time intervals/distances traveled/time elapsed. The underlying assumption is that the level of performance and its time of operation are related to the degradation of the lubricant. Thus, 100 km of unrestricted daytime high-speed driving on an autobahn in summer is assumed to degrade a lubricant more than 100 km of urban driving in autumn or spring. Thus, the aggregates of high power level operation over time are weighted more than the same period of low power operation. Integration of the high and low power level operation is already used in some vehicles to indicate to the operator when the system’s service is due and the lubricant must be replaced. The objective at the end of the service period must be that the lubricant still be “in grade,” therefore specification, and that the engine or machine not to have suffered “excessive wear” or component damage. This “state of grace” is readily achieved by the vast majority of lubricants in operational service through the development and testing of formulations. The major current development is for service intervals to increase in terms of hours operated or distance traveled. Thus, for light vehicles, service intervals are progressively increasing to 20,000, 30,000, and 50,000 km for light vehicles. A target of 400,000 km is envisaged for heavy duty diesel engines or their “off-road” equivalent.

29.1.2 The Effects of Deterioration Lubricants are formulated from a base oil mixture and an additive pack, as described elsewhere in this volume. The base oil is usually a mixture of base oil types and viscosities chosen for their physical and chemical properties and their costs. The additives form part of an additive pack to protect oxidation, wear, acidity and corrosion, to remove and disperse deposits, maintain a specified operating viscosity range, and minimize foaming. A filter in the lubricant circulation system should remove suspended particulates above a certain diameter. Lubricant degradation occurs throughout its service life and the baseline for change is reached when its further deterioration would lead to a level such that it cannot protect the system from further excessive wear. This occurs because the lubricant has become physically unsuitable for further service use for several

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29-5

separate or joint causes: • It has become too laden with particulate dirt. • It’s viscosity has increased/decreased beyond its specification limit. • It’s additive pack has become depleted in one or multiple components. Often the additive component actions are interdependent, thus oxidants may protect other additive actions. • Abrasive and corrosive materials can cause bearing damage, or bore polish by removing the crosshatched honing marks, which maintain the lubricant film, or in extreme cases, “scuffing” of piston and bore. These effects are often interdependent and will cause further changes either directly related or through catalytic effects. When these lubricant deterioration effects occur in such complex systems as lubricant formulations, then a structured approach is needed to understand and solve the problem.

29.1.3 Physical Causes of Deterioration A lubricant formulation becomes physically unsuitable for further continued service use through a range of the following causes: 1. Internal sources: internal contributing sources are those which are either introduced into a system by the production or repair process, as: (a) Textile materials such as (production line) cleaning cloths, contributing “lint,” which compacts into obstructions of oilways. (b) Metallic materials such as metallurgical cutting residues and welding repair particulates or production grinding processes, or by the operational process, of either fuel or oxidative use, as follows: • Harder/softer particulates from the partial oxidation of lubricants as harder particulates from longer, C30 hydrocarbons, as in lubricant hydrocarbons, and softer particulates from shorter, C15 , hydrocarbons, as in diesel fuels hydrocarbons. • Through defective sealing systems, which allow ingress of silicaceous abrasive sources. • Fuel condensing into the lubricant and reducing its viscosity, or together with condensed water, forming an emulsion of low lubricity value. Cooling water ingress into the lubricant system through defective seals is another source of water contamination. 2. External sources: external contributing sources, predominantly grit and dust, are those either introduced into a mechanical system by: (a) Infiltration through exhausted and inefficient oil filters (b) Filling through unclean filler pipes/tubes (c) Lubricant reservoirs open to the (unclean) atmosphere (d) Through overwhelmed air filters, as in desert area operations The debris of system wear, abrasive wear products from combustion processes, and defective sealing materials are physical causes of lubricant deterioration. Another obvious physical cause of degradation is to add an incompatible lubricant to an existing formulation in an existing system — while the base fluids may be miscible, their additive packs may be incompatible and precipitate (“drop out”), leaving the circulating fluid as a simple base oil system with little mechanical/tribological protection. In most cases, the physical causes of lubricant deterioration are simply related to good maintenance, or the lack of its meaningful application, simply put as “good housekeeping.”

29.1.4 The Effects of Lubricant Chemical Deterioration Of all the chemical causes of lubricant deterioration, oxidation is the most important. It has extensive onward connections to the formation of organic acids, usually carboxylic acids, sludges that lead to resins/varnishes, which in turn bond carbonaceous deposits onto system components. Oxidation forms

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hard carbon from heavy hydrocarbons such as lubricant base oils, engines become very dirty and if the oxidation is sufficiently severe, then essential small orifices such as filters, minor oilways, and crucial orifices such as undercrown cooling jets become blocked and rapidly cause severe wear problems. Oxidation is temperature dependent and as a chemical reaction, is subject to the Arrhenius effect of reaction rates doubling/trebling for every 10◦ C increase in temperature. Thus, a reaction rate of unity at 300◦ C will increase to between 2 and 3 at 310◦ C, to between 4 and 9 at 320◦ C, and between 8 and 27 at 330◦ C, and so on — a compound increase. This has important implications for trends in increasing engine power densities, smaller lubricant volumes, and reduced cooling effects due to vehicle aerodynamics, which lead to increased engine operating temperatures, including its lubricant system. Future lubricants must withstand higher operating temperatures using smaller volumes for longer service intervals. Advanced lubricant formulations must be developed, which can operate at consistently higher temperatures to prevent their deterioration below levels that protect power train systems for extended, longer, service changes. The reserve concentration of unused, effective antioxidant in the lubricant during its service life is a crucial factor. Exhaustion of the antioxidant in the continuous use of a lubricant rapidly leads to the mechanical deterioration of the system. It is not sufficiently appreciated that heavier hydrocarbons, as used in lubricant base oils, have up to 10% of air dissolved or entrapped within it, the difference is semantic. The mechanical movement of the lubricant, as flow, agitation, or foaming, will maintain the air/oxygen concentration in the oil and increase the rate of oxidation. High temperatures will also affect the base oil molecules and additives directly. Thermal degradation is selectively used in refineries to reform hydrocarbons at temperatures similar to those by lubricants experienced within engines. Under the relatively uncontrolled thermal breakdown conditions within an engine, base oil molecules can break down into smaller molecules, “cracking,” or become functionalized with carbonyl groups, particularly, and undergo polycondensation to form varnishes and gums, which trap and sequester carbonaceous particles. The thermal stability of base oils is an important parameter in their selection. Additives are destabilized by high engine operating temperatures, dependent upon the extent and duration of their exposure to these high temperatures within the engine system, such as the ring zone and valve guides. The term “additives” covers a wide range of compounds, which can contain sulfur, phosphorus, and chlorine. Complex additives can break down to form a range of smaller compounds; thus, Zinc Di-alkyl Di-thio Phosphates (ZDDPs), antioxidant and antiwear agents break down in the ring zone of diesel engines to form organic sulfides and phosphate esters [1]. But reaction between additives — additive interaction — caused by exposure to high temperatures, not only depletes those additives but can also generate sludge deposits. The intermediates may also be corrosive to the system. There are several overall tests for the antioxidant reserve/antioxidancy of an oil, new or used, as either the ASTM 943, 2272, and 4310 tests, also the IP 280 tests. Of these are the following: • The Rotating Bomb Oxidation Test (RBOT), ASTM 2272 where a rotating bomb is loaded with a lubricant oil charge, pressurized with oxygen in the presence of a copper catalyst and water within a glass vial. The time recorded for the oxygen to deplete, by reaction, and its pressure to fall by a specified increment of 25 psi (1.74 bar). This method is operator-intensive and has a range of random errors greater than the other. • Pressurized differential scanning calorimetry (PDSC) method, CEC-L85T-99-5 is a relatively lowcost test with much improved reproducibility, where a small (8 mg) sample within a very small cup is held under 35 atmospheres pressure of air in a differential scanning calorimeter at 190◦ C. The time for the overall additive function to be exhausted by the combination of high temperature and the diffusion controlled oxidizing atmosphere and the residual hydrocarbon combustion to give an exotherm, as in Figure 29.1, is the “induction time.” New lubricant formulations will have longer induction times, which will gradually reduce for used samples of that formulation as its service life proceeds. A “zero” value for an antioxidant “induction time” indicates that the lubricant sample is substantially degraded and unprotected against further, and substantial, oxidative attack.

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Enthalpy

Exotherm for new lubricant Exotherm for used lubricant Induction times

tused

tnew

Time

FIGURE 29.1 PDSC induction time plots for new and used lubricant samples.

Wear or system failure

Terminal wearout/ failure state Initial rapid decrease in wear or system failure – “break-in”/“running-in”

Normal wear/system life — low level of steady state wear/failures Onset of terminal wear/failures

Operating time

FIGURE 29.2 The “bath-tub” curve for wear/system failure.

29.1.5 The “Bath-Tub Curve” All systems wear but at different rates in their serviceable life. The pattern of wear is well described by the “bath-tub” curve, which is a plot of “wear” against time, Figure 29.2. It can also be regarded as a plot of system failure against time. A “bath-tub” curve does not describe “wear” (or “failure”) for individual systems but is a statistical description of the relative wear/failure rates of a product population with time. Individual unit can fail relatively early but with modern production methods, these should be minimal, others might last until wear-out, and some will fail during the relatively long period, typically called normal life. Failures during the initial period are always caused by material defects, design errors, or assembly problems. Normal life failures are normally considered to be random cases where “stress exceeds strength.” Terminal “wear-out” is a fact of life due either to fatigue or material depletion by wear, from this it is self-evident that the useful operating life of a product is limited by the component with the shortest life. The “bath-tub” curve is used as an illustration of the three main periods of system wear/failure, and only occasionally is wear and failure information brought together into a database and the initial, normal, and terminal phases of

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system wear failure measured and calibrated. The timescales for these phases usually vary between one system and another. However, when condition monitoring is used to monitor the wear of a system, then a gradually increasing level of iron in each sample taken at service interval lubricant changes indicates that an engine has entered the final phase of its service life and its replacement and overhaul is becoming due. The necessary replacement arrangements can be made without failure or unexpected interruption of service. This saves costs because the engine is worn, but not damaged, readily and economically overhauled, the operation is planned and service interruption is minimized. Informed replacement of worn systems or components is usually estimated to have a direct benefit/cost ratio of 10:1, rising to 20:1 when indirect costs of unexpected interruptions of service are included.

29.2 Field Tests for Lubricant Deterioration Laboratory analyses of lubricants are necessarily done in laboratories; they are accurate but delayed unless, unusually, an operating site has its own laboratory. There is a good case for simple field tests, which may be less accurate but gives an immediate indication. Often the operation is physically separated from a laboratory, as in a merchant or naval ship, and needs reliable, simple tests.

29.2.1 Direct Observation of Lubricant Condition An experienced observer of lubricant condition will give considerable attention to the color of a lubricant sample — it is helpful to compare with an unused sample. Oxidative and thermal breakdown of a lubricant, often beyond exhausting its antioxidant reserve, gives a darker, more brown, color. The deepening in color is also associated with a very characteristic “burnt” odor, which is recognizable when experienced. The viscosity of the sample will also increase.

29.2.2 Field Kits for Lubricant Condition Various “field kits” are available to measure the essentials of lubricant condition, such as viscosity, water content, particulates, and degree of oxidation. These were called “spot tests” in the past but have improved in reliability to be acceptable for continuing analyses where access to laboratory tests is limited, such as on ships or isolated sites. Viscosity is readily measured by using a simple “falling ball” tube viscometer in the field on site. Comparison with an identical apparatus, often in a “twin arrangement” containing a new sample of lubricant gives a direct comparison of whether the used lubricant viscosity has relatively increased or decreased by the respective times taken for the balls to descend in their tubes. The simplest method to determine particulate levels in a sample of a degraded lubricant is the blotter test, where a small volume of oil sample is pipetted onto a filter paper or some other absorbent material. This is generally known as the “Blotter Test,” which can take various forms, either using a standard filter paper or a thin layer chromatographic (TLC), plate. The measurement concerned is the optical density (OD) of the central black spot. The higher the level of particulate, the denser (darker) the spot. The assumption is that the spread of the lubricant sample disperses carbon particulate within an expanding circle and that the optical density of the carbonaceous deposit is a direct measurement of the mass of particulate present in that sample. The system can be quantified by use of a simple photometer, for fieldbased simple systems, or a spectroreflectometer for laboratory measurements. Methods of automating these types of systems have included the following: • Automated, accurate, constant volume pipetting of the oil samples • Video measurement of the oil sample blot on the filter paper, thus its “OD” • Data recording of these results Despite many attempts and applications, these advanced methods have not achieved universal acceptance, possibly because of the increased complications built onto an initially simple test. Another,

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Black soot ring White oil ring

Development over time

FIGURE 29.3 TLC plate soot spot/oil ring dispersancy test.

and major, problem is the heterogeneous nature of the samples presented for analysis, which give different responses, arising from: • Different base stocks, such as the differences introduced by the mineral, semisynthetic, and synthetic base stocks used in modern lubricant and hydraulic formulations. • Different formulations, such as the differences between hydraulic, automotive, aerospace, and marine fluid formulations, a high dispersancy oil spreading its carbonaceous matter over a greater area than a low dispersancy oil. Marine lubricant formulations are an interesting case to consider. The lubricant volumes used per engine/vessel are very large, of the order of 103 l. The fuel used is high in sulfur, not being controlled to the same extent as land-based automotive diesel fuel, causing extensive additive and base oil degradation. The general case is for vessels to pick up the available top-up lubricants whenever they dock in various ports, leading to heterogeneity of base stock and additive formulation. These factors lead to scatter in the particulate signal/concentration plot. A further development of the“Blotter Test”is to use TLC plates, which are more uniform than paper. The intensity of the black spot from a 50 µl aliquot can be measured and, if its image is captured electronically, may be integrated across its area. But the black carbonaceous spot will also have a base oil ring extending beyond it, seen either as a change in white shade or a fluorescent area under UV illumination, Figure 29.3. The diameter of the white oil ring measures the movement of the lubricant and the black soot ring the movement of the soot particulate. This can be developed into a measurement of dispersancy for the oil sample. Dispersancy is a difficult property to measure, analysis of the dispersant concentration may indicate the amount of free dispersant in the sample together with a variable amount of dispersant desorbed from the particulate, an unsatisfactory measurement. The most effective way to measure dispersancy is to measure the dispersancy ability of a sample, not the concentration of dispersant. The dispersancy of a sample can be measured by the ratio of the black soot ring to the white oil ring. While this is not absolute, the change in dispersancy over the course of an engine test or the service life of a lubricant can be followed by the change in the spot/ring ratio, as the CEC97-EL07 development method. The method is very reproducible, provided that all of the following are considered: • Multiple samples are taken, which is much easier than the previous methods. • The sample images are captured using high resolution optical electronic methods and the area of each spot integrated, as the edges of the spots are often uneven in detail. • Each micropipetted sample is accurately and reproducibly dispensed. The ratio of the “spot” diameters for the white oil ring and the particulate measures the ability of the lubricant sample to disperse carbon particulates, a high ratio indicating a high level of dispersancy remaining in the lubricant. Equally, a low ratio of carbonaceous black spot to the radius of the oil blot indicates a low level of dispersancy. Dilution of a used lubricant sample with a light hydrocarbon such as “Petroleum Ether 60/80” and subsequent filtration through a standard filter paper will indicate the nature of the larger particulate debris, emphasizing metal particulate debris. Microscopic examination of the metal debris can show the nature of the larger metallic debris, which indicates the pattern of wear.

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Water content can be measured “in the field” by mixing a lubricant sample with a carbide tablet in a sealed stainless steel bomb. The measurement of water content is through the reaction of the carbide tablet (or calcium hydride in an alternative model) to generate gas pressure within the bomb. The pressure level generated is a measure of the water content of the sample. An alternative quick test for water content is the “crackle test,” where a small lubricant sample is suddenly heated. This can either be done by suddenly inserting a hot soldering iron bit into the sample — if water is present, a “crackling” noise is heard, which is absent for dry samples (the noise comes from steam generation in the sample) or small drop of sample can be dropped from a syringe onto a “hot” laboratory hot plate, when again a “crackle” will be heard if the sample is “wet.” From experience, the limit of detection is taken to be 0.1% water. The degree of oxidation can be measured by a simple colorimeter using a standard sample to measure color, ASTM D1500. The trend compared to previous values is the important observation. If the change occurs early in the service of the lubricant, then the antioxidancy reserve of the lubricant is being rapidly depleted or the lubricant is being contaminated. It is important to consider the change in color in combination with values and changes determined for Acid Number and viscosity for the same samples. Other simple tests are available in addition to those described above, as a suite packaged into a portable package for measurement of lubricant degradation in isolated situations such as remote mines and onboard ships.

29.3 Laboratory Tests for Lubricant Deterioration Some introductory general remarks are useful: • Results from laboratory tests for lubricant deterioration are of much greater value if the original, virgin, unused, lubricant is used as a benchmark. • Similar tests apply to most forms of lubricants as the deterioration challenges they face are chemically and physically similar. • However, the results from similar tests for different forms of lubricants must be considered in the context of each lubricant’s application. The advantage of laboratory tests is that they should have a background of both quality assurance and control. From this, they have serious weight in solving problems, assessing oil change intervals, what preventive maintenance is required from condition monitoring to conserve the system, and as well lubricant resources. The primary objective of a laboratory analysis program for lubricant samples is to ensure that they are fit for further service. If the lubricant is unfit, or becoming unfit, for further service, then it must be replaced. The benchmark for a laboratory program of sample analyses to assess a lubricant’s deterioration is to offer a rapid turnaround for analytical results, their assessment against limit values, and reporting back to the client. Isolated heavy plant mining operations can have lubricant analytical sample reporting times of weeks due to transport and communication issues; intensive transport systems in developed countries can expect less than 24 h reporting, such that a sample taken 1 day will be analyzed and reported upon before the next day’s operation commences and the appropriate action taken. An equally important benchmark is for laboratory to meet the various national or international standards, such as the ISO 9000 series. The use of certified analytical standards and accredited solutions is part of a complete package, which best involves a collaborative program of regular analyses of samples sent from and collated by central standards body. All apparatus and substances used should have an audit trail for standards and calibrations that are maintained. This is not only a good practice but necessary to respond to any implied liabilities, which may arise later. Of the many tests available, the major issues of lubricant deterioration are addressed by analyses of viscosity and viscosity index (VI), trace metals, particulates, ash, acidity/base reserve, and water contents. Other minor issues are color, demulsibility, foaming, rust testing, infrared spectroscopy, and to a certain

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extent, x-ray fluorescence (XRF). Gas and liquid chromatographies, x-ray diffraction, interfacial tension, and density are peripheral techniques that might be used to investigate unusual occurrences.

29.3.1 Viscosity and Viscosity Index Viscosity is the foremost quality of a lubricant to be measured. A lubricant must maintain its viscosity to effectively protect a system against seizure. Variations in viscosity are usually associated with effects, which show up in other analyses; therefore, a multidimensional approach is needed to consider the root cause of the change. Viscosities of lubricant samples are now measured by automated systems, taking samples from multiple sample trays, either circular or linear, and injecting them into either kinematic or absolute viscometers thermostated at either 40 or 100◦ C. If separate viscometers at these temperatures are used, then the VI of the sample can be calculated. Standards are inserted into the flow of samples through the system for quality control. Individual measurement using manual stopwatches and U-tube suspended viscometers are now rarely used in laboratories. A small increase in lubricant viscosity may be due to evaporation of the lighter ends of the base oil after prolonged high level operation. Beyond that, significant increases in viscosity, up to 10/20% being regarded as severe, result from the inadvertent replenishment with a higher viscosity lubricant, extensive particulate contamination, and extensive base oil oxidation. The particulate contamination as well as extensive oil oxidation will be readily seen, the latter on its own as an increasing dark brown coloration. The black particulate contamination will obscure the brown oxidation color. Oxidation effects will also appear in the Fourier Transform Infrared (FTIR) spectra and decrease in the PDSC antioxidant reserve time. A decrease in the viscosity of operating engines is usually due to fuel dilution, a characteristic occurrence when an engine idles for a prolonged period. A locomotive used for weekend track maintenance train duties will run its diesel engine at idle for periods of several hours and its lubricant will show a significantly decreased viscosity afterward due to fuel dilution. If subsequently used for normal, higher power duties, the increased lubricant temperature will evaporate the condensed fuel and the viscosity returns to its previous value. A more serious occurrence is when fuel and water are extensively condensed in the crankcase of a very cold engine at start-up. During short journeys, when the engine lubricant rarely becomes warm enough to evaporate the condensed fuel and water, the two contaminants can combine to cause the additive package to precipitate out from the lubricant formulation. The engine may then have “oil” but is then left with considerably reduced protection wear. Measuring fuel dilution in diesel lubricants is difficult and is discussed later in the subsection on “flash point.” Fuel condensed into a lubricant has the role of a solvent and the same effect of decreased viscosity is found when a solvent becomes entrained, such as a refrigerant fluid. Chlorofluorocarbons (CFCs) are well on their way to removal and nonreplacement from refrigeration systems but their replacements, the hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs) have the same effect of reducing viscosity if allowed to leak through seals or rings and dissolve in a lubricant. Viscosity index improvers (VIIs), are long chain polymers of various basic units. Their different structures resist high rates of mechanical shear, as in bearings or in the ring pack/bore wall interface, to different extents. While there is a separate effect of temporary viscosity shear loss, lubricants with VIIs can suffer permanent viscosity shear loss due to breaking of the polymer chains. The initial lubricant selection process should have considered how robust the formulation was to permanent shear thinning. Tests for this include high temperature and high shear procedures such as ASTM D4683 and D4741. If the viscosity of a lubricant changes during its service use then its VI, will change necessarily. The major cause of a reduced VI is caused by breaking of some of the polymeric VII polymer molecules to give smaller chains of less effect. There are two effects — reduction in the molecular weight of the VII additive will reduce the viscosity of the lubricant formulation at both 40 and 100◦ C and also reduce the temperature related VII effect. The latter effect normally has the greater weight so that the permanent

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shear breakdown of polymeric VII additives reduces the lubricant’s VI. It is not unknown but rare for the VI of a used lubricant to increase in service use, often associated with extensive oxidation.

29.3.2 Trace Metals The term “trace metals” in a lubricant sample not only covers metals generated by wear in the system but also the elements from the additive pack. While the determination of trace metals for a “one-off ” sample gives some insight into the condition of a lubricant, the major value of trace metal determination lies with long-term condition monitoring. The “bath tub” curve of Figure 29.2 is recalled here — following the level of iron fine particulate in a series of regularly sampled lubricant from a system is an essential part of condition monitoring. The “break-in” or “running-in” phase, normal wear, and the gradual increase in wear element determination can be followed running over many hours and lubricant service changes. The onset of terminal wear can be detected and followed, with arrangements put in place to remove and replace the engine system. Levels of wear elements measured are usually iron (from bores and crankshafts), lead and copper from bearings, aluminum from pistons, and chromium from plating on piston rings. Others may be added to follow specific effects, for example, sodium levels indicate the ingress of cooling water and its additives, silicon levels indicate the ingress of sand and rock dust. It is important to recognize that the level of wear elements in a system’s lubricant is individual both to system design and to individual systems. Thus, levels of iron in the normal wear phase of engines will be different from one design to another; in addition, there will be some variation between the normal wear phase iron levels of engines of the same design. The quality of the lubricant used will also affect the level of wear metal, the higher the quality of lubricant, the lower the level of wear elements. The emphasis for assessing the condition of lubricated systems is placed upon the trend in wear element levels. While the iron level in the lubricant of one engine may be higher than another, it is the trend for successive samples over time in the measured levels, which is important. Wear processes in lubricated systems rarely occur for one metal. Increases in the levels of several wear metals can indicate the occurrence a particular wear process or contamination. Table 29.1 describes wear elements found in lubricants in service life. Wear metal analyses have additional effectiveness when combinations of enhanced element rates are considered, such as for a diesel engine. Combinations of enhanced wear elements are unique to each operating system design and its pattern of use. “Expert systems” applied to an extensive data system can be used to develop “rules,” which indicate which main assemblies or subassemblies are developing enhanced rates of wear and require attention for certain engine designs. The examples given in Table 29.2 are typical for certain applications — other systems may have different combinations for wear patterns, it is for the expert system to recognize them. More extensive combinations of elements indicating particular wear patterns by system components can be developed, such as using the “principal indicator” and associated “secondary indicator” elements. Cost-benefit analyses of spectroscopic oil analysis programs, with the acronym “SOAP,” have been demonstrated in many applications to be very significant. Continuously and heavily used plant, such as diesel express trains, where daily oil sampling and analysis gives an immediate cost-benefit ratio of 10 : 1 in direct costs and 20 : 1 for indirect costs when service reliability benefits are included. The analytical methods for wear metals have generally moved to inductively coupled plasma (ICP) atomic emission systems. A small sample is automatically extracted from a sampling bottle, diluted with kerosene and sprayed into the ICP analyzer plasma torch at 6000–8000◦ C. The very high temperature of the plasma excites the metal particulates to high energies, which emit light of a characteristic atomic wavelength. Duplicates (or more) are readily programmed. The emission from each metal present is detected and reported, the cost of additional wear element detection is marginal once the ICP system is set up. The ready availability of duplicate sample determinations and insertion of calibration standards gives a high level of quality control as precision, accuracy, and reproducibility to the final results. The analytical data generated by the ICP system is readily handled, quantified, and then placed into a file for that engine

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TABLE 29.1 Wear Elements in Lubricants and Their Sources Source Major elements aluminum Chromium Copper

Iron

Lead Molybdenum Nickel Silver Tin Titanium Zinc Minor elements Antimony Boron Magnesium Manganese Sodium Silicon

Primary component of piston alloys, also bearings, washers/shims and casings of accessories. From corrosion of engine blocks, fittings, and attachments Used as a hard(er) coating to reduce wear, indicates wear of chromium plating on engine bores, shafts, piston ring faces, some bearings and seals With zinc in brass alloys and tin in bronze alloy wearing components, copper present in journal, thrust, and turbocharger bearings, also cam, rocker, gear, valve, and small-end bushings. Also, fabricate oil cooler cores Still a major, massive component of engines, gearboxes, and hydraulic systems. Lubricant contact through cast bores, cylinder liners, piston ring packs, valve guides, rolling element bearings, chains, and gears. Difficult decision given by wearing component increased trace levels of iron In bearings, solder joints as “lead/tin alloy” and also seals A wear reduction coating on first piston ring faces for some diesel engines From valves, turbine blades, turbocharger cam plates, and bearings Alloys in bearings, bearing cages, and bushings for diesel engine small ends, turbochargers and rolling element bearing applications in gas turbines Common alloy in bearings with aluminum, bronze and brass fittings, seals, and also in cooler matrix solder Top end of market, gas turbine bearing hubs, turbine blades, and compressors With lead and tin in common alloys such as brass and also some seals May be used in bearing alloys Borates used as cooling system anticorrosion agents, presence in lubricant and hydraulic fluids shows leak in cooling system matrix Increasingly used as an alloy with aluminum for accessories and casings From corrosion of manganese steel alloys, occasionally in valves Usually sodium borate as cooling system anticorrosion agent. Increasing trace presence in fluids shows leak in cooling system matrix, marine applications indicate ingress of coolant sea water Piston wear. As silica, indicates road dust ingress, particularly damaging as hard particulate, which causes high levels of wear, shows air filter and breather system failure, particularly mining and deserts

TABLE 29.2 Elements

Some Indicative Combinations of Wear

Elements Sodium and Boron Lead and Copper Copper, Silver, and Iron Chromium and Iron Silver, Copper, and Lead Iron and Copper

Indicative cause Coolant leakage into lubricant, as through head gasket failure Main or big-end bearings Turbocharger bearings Piston rings Small-end bush Oil pump wear

system, which can then be compared with previous results. This is concentration level “trending” in its simplest form. The overall effect is to give a high throughput of high quality analyses at low cost. While the automated sampling ICP multiple element system has a high capital cost, of £150–200 k ($300–400 k) each, the high sample throughput can cut the unit cost per sample down to 50 p ($1). An atomic absorption (AA) apparatus can be used instead of the ICP system but suffers from the disadvantage of only determining one element per analysis from the nature of this method. The older emission system of an electric discharge between either still or rotating (“Rotrode”) carbon electrodes is still used but the advantages of the ICP system for high throughput of samples are gradually displacing it. The ICP spectroscopic technique and oil samples are brought together as a condition monitoring system.

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Handbook of Lubrication and Tribology TABLE 29.3

Corrective Levels for Lubricant Deterioration

Deterioration level Normal Alert Urgent Hazardous Dangerous

Action Within average, no action Within average ±2σ , action → increase sampling frequency Within average ±4σ , action → maintenance needed, can be deferred Beyond average ±4σ , action → immediate maintenance, no deferral. Or trend in analysis >60% average Trend in analyses up to 90% of alert level, action → shutdown/recall immediately/immediate urgent maintenance

It is meaningful to analyze trends in the wear element test data, which monitors the deterioration of the oil condition. Absolute and rate of change data concentration values can be used to assess the deterioration of a lubricant or hydraulic fluid — the ideal scheme, with regular sampling, servicing, and replenishment at preprogrammed intervals. It is rare for this regularity to hold; the reality is that sampling/servicing and replenishment of fluids occur irregularly and this must be adjusted numerically in the trend data. From these “trending analyses,” element concentration indicators can be developed by various statistical methods using system failure modes to set individual wear metal levels at which corrective or remedial measures must be taken for the deterioration of the lubricant, such as in Table 29.3.

29.3.3 Particulates and Ash in Lubricants The accurate measurement of particulates and ash in a lubricant sample is very important in assessing its deterioration for the excessive build-up of soot, dirt, or particulates in general can prevent the normal protective function of that lubricant. The term “particulates” covers a wide range, including insoluble matter, sediments, and trace metals as very fine diameter particulate. Larger metal particles such as metal flakes and spalled debris are not covered, these being covered by separate analyses and filtration. 29.3.3.1 Dirt and Particulates in Lubricants Controlling the cleanliness of any lubricant or hydraulic system as it deteriorated with use was very important in the past and will be even more important in the future, because of the following reasons: • System reliability is increasingly important and a major contributor to equipment failure is particulate contamination in the system operating fluid. • Systems perform at higher energy levels for longer periods and maintained to be “cleaner” so as to deliver that performance. • Equipment tolerances are decreasing for high precision components (∼5 µm clearance or less) and in automotive and hydraulic components they are increasingly common. Smaller particulates, for example, 2 µm dependent upon its nature, can agglomerate and clog sensitive components such as control and servo valves. • For automotive applications, two trends lead to increased particulate levels: Exhaust gas recirculation, for environmental exhaust emission reduction, primarily for NOx, having the additional beneficial effect for emissions of depositing particulate into the lubricant rather than being emitted. However, this creates a problem of enhanced particulate levels for the lubricant. Strong consumer pressure for increased service intervals, already up to and beyond 50,000 mi (80k km) for trucks and 30k mi (∼50k km, or every 2 yr) for some new 2005 light vehicles. Lubricant must last longer and yet meet enhanced performance standards. Enhanced levels of particulate are now envisaged, well above 1%, up to 2 or 3%, a steep challenge for the lubricant to remain effective under these conditions.

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29.3.3.2 Useful Definitions “Particulates” and “dirt” are descriptions, which require a more precise description and definition, as follows: • “Particulates” are small, up to 15 µm maximum, either carbonaceous, inorganic compounds or fine metal particles, where the metal particulates result from “rubbing wear.” • “Dirt” is road dirt ingested by faulty induction air filters, poor seals or defective/absent air breather components; the parts that survive are usually hard particles such as silicates (from sand, etc.). • “Metal debris” is comprised of larger metal flakes or spalled particulates resulting from catastrophic micro-failures or incipient major failures such as parts of gear teeth being separated. Hydraulic fluids develop haze or very light deposits over a considerable time of their service life, petrol/gasoline engines develop black particulates slowly over their service life, while diesel engines rapidly develop black particulates. The operating limit of circulating lubricant filters is in the range 10 to 15 µm whereas it has been shown that removal of the “larger,” >10 m particles from a circulating lubricant system can reduce catastrophic bearing failures by 25%. Further, for hydraulic systems it is claimed that 80% of failures can be avoided if particulates >5 µm are removed by filtration. While not going to these levels of filtration, higher levels of filter efficiency are now incorporated into new designs. This must happen to meet the enhanced levels of filtration required over the enhanced periods of service operation. However, one problem is that the enhanced levels of filtration can remove the small, fine, metal particulate, which is used for wear data and trend analysis. 29.3.3.3 Particulate Analyses There are a number of measurements available for the measurement of soot in lubricants. These measurement methods can be grouped into three categories, as where the particulate is: 1. Removed from the liquid, then oxidized while measuring the mass loss. 2. Separated by addition of solvents to the lubricant sample and the precipitated mass measured. 3. Measured within the neat, or diluted, lubricant sample for absorbance, scatter, or obscuration at a given wavelength. 29.3.3.4 The Enhanced Thermogravimetric Analysis, ASTM D5967 Appendix 4 (Colloquially Known as the “Detroit Diesel Soot Test”) Total particulate in a degraded oil sample is determined by thermogravimetric analysis (TGA), where 20 mg of oil in a pan on one arm of an electronic balance is heated under programmed temperature furnace environment in a nitrogen atmosphere. Differentiation is made between carbonaceous and incombustible ash by increasing the temperature and changing to an oxygen atmosphere. A 20 mg sample is larger than normal but is necessary because the final objective, the soot content, will be less than 1 mg. The temperature environment is held at 50◦ C for 1 min, raised to 550◦ C at a rate of 100◦ C/min, maintained isothermally for 1 min, and then raised to 650◦ C at 20◦ C/min. The method considers the residual sample at this stage to be composed of soot and incombustible material with liquid hydrocarbons removed. The atmosphere is then switched to oxygen and the furnace temperature raised to 750◦ C at 20◦ C/min and maintained for a stable weight for at least 5 min. The changes in weights at different temperatures and atmospheres are due to soot being the difference in weight between 650◦ C in nitrogen and 750◦ C in oxygen. The residual material is incombustible ash and metallic residues, assuming that all of the remaining lubricant base stock is driven off and oxidized at the higher temperatures under oxidizing conditions. 29.3.3.5 Optical Particulate Measurements A very desirable feature in particulate measurement is a linear relationship between particulate signal, by light absorption or scattering, and particulate concentration. This relationship generally holds as a linear relationship of a certain slope up to ∼1.5% particulate concentration, followed by a linear relationship with a higher slope at higher particulate levels. While an overall linear relationship is very desirable,

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Absorbance at 500 nm

1.0

CEC L-82-A-497 Optical particulate measurement sample dissolved in toluene

0.5

Particulate concentration

FIGURE 29.4 toluene.

CEC L-82-A-497 calibration plot for particulate determination in degraded lubricant by dilution in

the major problem is the change in relationship between signal, however derived, and particulate concentration in the region of 1.5% concentration. Two methods measure sample particulate concentrations, one infrared by direct sample absorption and one in the visible by dilution in toluene. The visible method, capillary electrophoresis chromatography (CEC) L-82-A-497, “Optical Particulate Measurement” dilutes the degraded oil sample in toluene, a solvent, which disperses all of the particulate, and then measures the absorbance of the diluted solution at 500 nm in a spectrophotometer. Standardization uses a lubricant or hydraulic fluid sample of known pentane insolubles content to construct a calibration curve, Figure 29.4. The method is quick, repeatable, and accurate, provided that the sample disperses well and does not cause light scattering, which will add to the apparent OD. This method was adopted by the CEC to measure soot developed in lubricant samples from the Peugeot XUD11BTE engine test and uses 0.1 g of oil sample in known aliquots of toluene. The solvent aliquot volume is increased to bring the OD within an acceptable range. The OD plot for lubricant samples dispersed in toluene and measured at 500 nm should be linear with a high correlation coefficient. The only drawback is that some additives or degradation products may cause light scattering and an incorrect result. 29.3.3.6 Infrared Measurements at 2000 cm−1 Soot does not absorb in any specific region of the infrared region but as small particulate scatters the incident radiation in a nonphotometric manner. Theoretically, light scattering of a spherical, uniform diameter, particulate is proportional to the fourth power of the wave number. From this, the background scatter in the infrared spectrum of a used lubricant containing particulate should decrease across the infrared region from 4000 down to 400 cm−1 . Background scatter does decrease but not as much as predicted by theory, probably because the particulate is not monodisperse and certainly not spherical. 2000 cm−1 is the chosen measurement point because there are no absorbing groups present in lubricants. Increase in lubricant absorption at 2000 cm−1 with engine run time are mainly dependent upon the mass of soot particulate present, with second order effects due to the effective particulate size and shape, therefore somewhat dependent upon engine type. High levels of soot particulate give high absorbance levels and inaccuracies in spectrophotometry, which can be overcome by using thinner path length cells. The results are in absorbance and need calibration for percentage soot. The advantage of the method is that it is a direct measurement on the sample, without the effects of adding solvents, the like, and that it arises from infrared measurements, which could be undertaken for another set of measurements in any case. The disadvantage of the method is that the sample spectra need to be the difference spectra, that is, the difference between the engine test run samples and the original, fresh oil, which may not always be available.

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29.3.3.7 Particle Size Distribution A more fundamental view of the nature of particulates in degraded lubricant and hydraulic fluids is the distribution of particulate sizes. This can be either done continuously, by light scattering, or discontinuously by a range of physical filters. The latter is self-explanatory and the first needs explanation. Particles suspended in a medium scatter incident light at an angle dependent upon particle size and also upon the wavelength of the incident light. The second is simplified by using a monochromatic source such as a laser. The particles are assumed to be spherical, a very broad-brush approximation. A variable correction factor is needed for the nonspherical nature of the particulates, such as a rodlike nature with a defined length/width ratio. Particulate light scattering optics uses a collimated laser light source, usually a He/Ne (red) laser, expanded by a lens into a broad beam, which diffuses the sample cell. The light is dispersed/scattered by the suspended particulate in the sample cell and then collected by a similar second lens and focused onto a detection plate. The detection plate samples the intensity of the scattered light at a large number of points and transformed into a particle size distribution by suitable software. The resolution of the method depends upon the spatial discrimination of the detector plate. Particle size distributions for a range of samples from engine runs using a range of related lubricant formulations show that these particulate distributions are interdependent, the smallest particulate size distribution leading to the successive growth of the larger particle size distributions. The interdependence of these particulate distributions measures the effectiveness of dispersants, for the particulate can successively agglomerate from the initial size of around 0.1 µm diameter to 1 to 7 to 35 µm and then larger diameters. If the dispersant within the lubricant is not degraded then the agglomeration process will be stopped or reduced. 29.3.3.8 Particulates in Hydraulic Fluids Hydraulic fluid cleanliness is crucial to the continued operation of hydraulic systems, avoiding component damage and failure. The level of cleanliness is many orders of magnitude down (better) from that accepted for lubricants. Instead of values of mass particulate, the emphasis for hydraulic fluids is on the number of particulates in the range of 2 to 15 µm, a range correlated to the probability of component problems. With this stimulus, several methods of electronic particle number counting have been developed, based upon the following: • Light absorption • Flow decay • Mesh obscuration These methods are continuous and easy to use; their main problem is the large amount of data that they generate for the size and number of particulates but without reference to the composition of those particulates. Wear metal or chemical analytical data is required to properly understand the complete picture of particulate composition in hydraulic fluids. A fundamental problem is the lack of suitable, repeatable reference standards. When used for equipment monitoring, it is very important that the response of the counter has a high particle size correlation with the size of particles, which cause damage to the fine tolerance components of the system. 5 µm diameter was regarded as the lower limit of damaging particles until recently, but this is now reduced to 2 µm as an indicator of potential damaging conditions, approaching the limit of discrimination between two such particles. One type of mesh obscuration particle counter uses three successive micro-screens of 15, 5, and 2 µm pore size, Figure 29.5. Laminar fluid flow through this array of screens generates pressure drops, caused by oversized particles partially blocking the respective pore size filter, recorded by differential pressure transducers. Count data from hydraulic samples is statistically derived through correlation with data from a calibration standard. This counter is effective for most oils of different levels of obscuration (lightblack) and is relatively insensitive to other counter-indicators such as entrained water and air in degraded lubricant samples. Another method of electronic particle size counting uses the blocking behavior of a particle size distribution in a degraded lubricant sample passing through a single, monosized micro sieve of either

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Handbook of Lubrication and Tribology Laminar flow

15 mm screen

5 mm screen

2 mm screen

FIGURE 29.5 The principle of the mesh obscuration particle counter. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). With permission.

Laminar flow

Micro-screen

Plunger

FIGURE 29.6 The principle of the flow decay particle counter. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). With permission.

15, 10, or 5 µm pore size, Figure 29.6. A correlation is assumed between the particle size distribution of an unknown sample and that of a standard. The measured parameter is the differential flow across the micro-screen, which converts flow decay measurements to an ISO cleanliness code. An optical particle counting method uses a path of collimated light passed through a hydraulic oil sample and then detected by an electrical sensor. When an translucent sample passes through the sample then a change in electrical signal occurs. This is analyzed against a calibration standard to generate a particle size and count database, linked to an ISO cleanliness value. Light absorption particle counter’s output values are badly affected by the following factors: • The opacity of the fluid raising the background value to the level that the instrument no longer works, overcome by sample dilution with a clear fluid. • Entrained air bubbles within the sample are counted as particles, which confuse the system, and are removed by ultrasonics and vacuum treatment. • Water contamination is more difficult to deal with, causing increased light scattering. But significant levels or water, such as >0.1 or >0.2% levels, will fail the oil anyway. The continued monitoring of particle cleanliness in hydraulic fluids within systems is a very important process to maintain the integrity and performance of complex hydraulic systems. 29.3.3.9 Ash Content The “Sulfated Ash” content of a lubricant is an important property and can be included under particulates in degraded lubricants. It gives a meaningful indication of the detergent additive content and is useful as a control test in the oil blending process. While it is a property only normally used for new

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formulations, results for degraded lubricants have considerable interference from both wear metals and other contaminants. The problem with sulfated ash arises from inorganic compound deposits in the ring zone and on the piston crown. The problem becomes very important when extensive deposits build up on the piston crown from low/medium power level operation, such as for a taxi engine in town. However, when such an engine is used at extended higher energy power levels, such as extended motorway journeys, the deposits on the piston crown become very hot, retaining heat and glow. They can become so hot that they melt part of the piston crown to the extent of penetration, that is, a hole, causing catastrophic deterioration of the engine, which is the downside of sulfated ash content. The upside of metallic detergent inclusion into lubricant formulations is their ability to reduce the deposition of carbonaceous substances and sludges in the ring zone and piston crown. The essence of the problem is to balance the level of metallic soap sulfonate in the original formulation and the amount of sulfated ash that results. Sulfated ash is a major contribution to the overall formation of ash, contributing to crown land deposits above the piston rings, valve seat deposits (and thus leakage through seat burning), and combustion chamber deposits. These deposits cause preignition of the gasoline/air mixture, leading to a decreased fuel octane rating for the same engine called octane rating decrease (ORD). It is beneficial to reduce the impact of this effect by minimizing ash deposits. Ash formed from lubricants can also contribute to diesel engine particulate emissions. Recalling that the sulfated ash content is important for new lubricants, the simplest test is the ASTM D842 Ash Test where the ash content of a lubricant is determined as a weighed sample, to constant weight, of oil burned for 10 min at 800◦ C. The mass measured is that of the incombustible solids, be they wear metals or other incombustibles such as fine metallic particles or silicaceous dust. The ASTM D874 Ash Test is an improved ASTM D842 method in that the oil sample is combusted until the carbon residue and metallic ash is left. Sulfuric acid is added, the sample is reheated and weighed to constant values. The last stage converts any zinc sulfate to zinc oxide. The sulfated ash tests indicate the concentration of the metal-based additives in fresh lubricant blends. Problems arise from (i) any phosphorus present forming pyrophospates of variable composition, giving higher and more variable results and (ii) magnesium sulfate being variably converted to its oxide. Carefully conducted, the sulfated test gives a reasonable measure of additive metals present in a lubricant formulation. The weight of metal present can be converted to the expected sulfated ash content by the conversion factors given below: To Estimate Sulfated Ash Content from Metal Content: Metal Conversion Factor — Metal % to Sulfated Ash Zinc Sodium Magnesium Calcium Barium

1.25 3.1 4.5 3.4 1.7

If the lubricant has been formulated with magnesium-based detergents or boron-based dispersants, then these methods of sulfated ash are unreliable. The sulfated ash test is also unreliable for used lubricants, due to the following reasons: • The presence of incombustible contaminants. • Additives will be degraded during service life and are thus changed chemically but the constituents will continue to appear in the ash residue at the same concentrations as for the new oil. • A trend toward ashless detergents, which undermines the relevance of the sulfated ash test as a measure of detergent in a formulation. It is important to check the sulfated ash method against reference blends wherever possible.

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Handbook of Lubrication and Tribology

29.3.4 Acidity and Base Reserve Determining the alkaline reserve or acid content of a degraded lubricant fluid should be straightforward by analogy to acid/base titrations in water. But this is the simplistic point that causes so many problems with determining “Base” and “Acid” numbers in degraded lubricant and hydraulic fluids. To thoroughly understand “Base Number,” an appreciation is needed to determine the following: • • • •

How it arises How it has been, and is currently, measured The problems of those analyses What this means for lubricant use/extended use and condition monitoring

While the idea of a “number” is simplistic and therefore appealing, the reality is complex and we need to look at the points made above, in order. 29.3.4.1 The Need for Base Number Measurement The need to measure the “Base Number” in some form as a property of a lubricant/degraded lubricant arises from the acidic products formed during the service life of that lubricant. The acid formation process can be rapid or slow, according to the stress that the lubricant is exposed to. The emphasis must be on the effect that the “service life” of the lubricant involves, in terms of either high temperature and pressure or over a short and intense, or a very long-term and less severe, service interval. The starting position is that most lubricant base fluids have some, maybe greater or lesser, basic properties that neutralize acidic components introduced into them. As the performance requirements of lubricants developed, it became evident that the naturally occurring antiacidic properties of unmodified base stocks were not sufficient to prevent lubricant and hydraulic oils becoming acidic and corroding the components of the system. The development of detergent additives had two effects: • The organic nature of the additives themselves had an additional, but marginal, antiacid contribution. • However, more importantly, the detergent additives had the ability to solubilize as inverse micelles alkaline, inorganic material such as calcium oxide/carbonate or the corresponding magnesium salts (much less used). These compounds react with acidic products formed in the lubricant to produce neutral salts, which bind the acidity as an innocuous product. Barium compounds are not used now because of toxicity problems. 29.3.4.2 Sources of Acidity-Induced Degradation Acidity in lubricants arises from two sources: • The (declining) sulfur content of fuels, forming sulfur oxides, primarily sulfur dioxide, SO2 . • The reaction (“fixation”) of atmospheric nitrogen by reaction with atmospheric oxygen in the high temperatures, 2000 to 3000◦ C, of the combustion flame front, forming nitrogen oxides such as NO, nitric oxide, and nitrogen dioxide, NO2 , primarily the former, which then slowly oxidizes to the dioxide. Sulfur and nitrogen dioxides, SO2 and NO2 , dissolve in any water present to give the mineral acids of sulfurous/sulfuric and nitrous/nitric acids. The two forms of each acid are given because the dioxides initially dissolve in water to give the first, weaker, acid and then oxidize to the stronger, second acid. Organic acids are formed by the partial oxidation of hydrocarbons. Normally, hydrocarbon oxidation is considered as going through to complete combustion with water and carbon dioxide as the final products. But combustion/thermal degradation can be partial, with hydrocarbon end groups forming carbonyl

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groups to make aldehydes, ketones, and carboxylic acids, the last as: R—C=O | OH Organic acids are not normally regarded as strong acids; acetic acid has a dissociation constant in water of 1.8×10−5 at 298 K and is regarded as a weak acid, the prime constituent of cooking vinegar. But various R-group substituents can increase the dissociation to make the acid stronger, such as for trichloroacetic acid. Two points to particularly consider for the strength of organic acids: 1. Acid dissociation constants increase with temperature, the higher the temperature, the stronger the acid. 2. The value given is for acetic acid in water. Acid : base interactions and equilibria are considerably different in other solvents, often making organic acids stronger. Applying these to organic acids in degraded lubricants, the lubricant is a drastically different solvent to water, which also operates at high temperatures. As an example of the strength of organic acids, the railways originally lubricated their steam engine cylinders with animal fats before hydrocarbon oils were available. The high steam temperatures within the cylinders degraded the fats into their constituent organic acids, which corroded the metals present, particularly the nonferrous metals such as copper, lead, zinc, and so on. The acidity generated within a degraded lubricant during its service life is a mixture of inorganic strong acids and weaker organic acids. This mixture is one of the causes of the analytical problems in determining the acidity of both the acids in, and the remaining alkaline reserve added to neutralize that acidity, in a lubricant formulation. This is the need to determine the Base Number in a lubricant, both new and used. It is a standard analytical measurement for degraded lubricants. 29.3.4.3 Measurement of Base Number An acid is normally associated with the bitter, corrosive, sometimes fuming in their concentrated form, properties of the mineral acids, classically sulfuric, nitric, hydrochloric, and phosphorous acids. There are others but these are the common mineral acids. Their common property is the ability to donate/give − a proton, H+ , to a base. Sulfuric acid then becomes an anion, such as sulfate, SO2− 4 , nitrate, NO3 , 3− − chloride, Cl , or phosphate, PO4 . The common bases as alkalis, such as sodium hydroxide, caustic soda, potassium, and ammonium hydroxides are strong bases with sodium carbonate as a mild alkali or weak base. Again, as for the acids, there are many others but these are the commonly used alkalis. The common feature of alkalis is the hydroxide group, OH− , which accepts the proton from the acid to form water, H2 O. Aqueous acids and bases in equal amounts neutralize each other to form a neutral salt and water, as in the standard neutralization of hydrochloric acid by sodium hydroxide: HCl + NaOH → NaCl + H2 O Whichever way this is done, by adding acid to alkali or the reverse, for equal amounts of acid and alkali, the end result is a neutral solution of pH 7. If the strength of one of the solutions is accurately known, then the concentration of the other solution can be calculated — basic chemical laboratory work. Neutralization is shown by an indicator with different colors in acid or alkaline solution, neutralization being shown by a color balance between the two forms. Litmus is one example of a neutralization indicator, being blue in alkaline and red in acid solution. Progress of acid/base titrations can equally be followed by other methods, such as: • The pH electrode combined with the standard calomel electrode to follow either the solution pH or the potential difference in millivolts, mV, between the electrodes.

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Handbook of Lubrication and Tribology

• The electrical conductivity of the solution between two platinum plate electrodes, because both the proton, H+ , and the hydroxide ion, OH− , have high conductivities relative to other ions and both H+ and OH− are at a minimum at the end point, pH 7. From these fundamental considerations, if the alkaline reserve (Base Number) of a degraded lubricant is a base, then it should be possible to titrate it against a standard acid solution to determine how much base is present. That is, the basis of Base Number determination, transferred over from water-based acid/alkali neutralizations to the analysis of new and degraded lubricants in a variety of organic solvent mixtures. Many acids have been used to titrate the alkaline reserve in a lubricant but they give different values, particularly for heavily used samples. 29.3.4.4 IP 177/ASTM D664 — Base Number By Hydrochloric Acid Titration This is a joint method developed by the Institute of Petroleum in the United Kingdom and ASTM in the United States and was the earliest methods for measuring the base content of a new or degraded lubricant or hydraulic fluids. It is still preferred by some operators and has essentially been reintroduced by the IP 400 method; see Section 29.3.3.7 later, with the same solvent and acid titration system but with a different detection system. The solvent for the titration of the lubricant/hydraulic sample must dissolve the sample and be compatible with the titrating acid. In this case, it is a mixture of toluene, isopropyl alcohol and a very small amount of water. The acid is dissolved in alcohol and the two solvents are completely miscible. The progress of the neutralization reaction is followed using a combination of a glass electrode and the standard calomel electrode, a standard nonaqueous solvent analytical procedure. The signal used is the potential difference between the electrodes expressed as mV. The neutralization works well for new and slightly used lubricants. The mV difference signal varies as a sharp sigmoidal form when mV is plotted against acid titration volume, Figure 29.7. The neutralization endpoint is at the mid-point of the sharp rise, as indicated. There is no problem with the analysis for new and lightly degraded samples, the neutralization curve is sharp, and the endpoint is clear. Problems arise as more extensively degraded lubricants are analyzed. The clear form of the neutralization curve slowly degrades with increased degradation of the lubricant sample until its form is lost and there is no clear endpoint, Figure 29.8. A procedure is suggested where an endpoint value to work to is used instead, but this is an unsatisfactory solution.

0.700

0.200

0.600

0.160

0.120 Endpoint

dE/ V

dE/ V

0.500

0.400

0.080

0.300

0.040

0.200 0.00

2.00

4.00

6.00

8.00

0.000 10.00

V/ml

FIGURE 29.7 method.

mV vs. volume plot for the titration of new/slightly degraded lubricants by the IP 177/ASTM D664

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0.400

0.060

0.340

0.048

0.280

0.036 dE/ V

dE/ V

The Degradation of Lubricants in Service Use

Endpoint

0.220

0.024

0.160

0.100 0.00

0.012

2.00

4.00

6.00

8.00

0.000 10.00

V/ml

FIGURE 29.8

mV vs. volume plot for the titration of heavily degraded lubricants by the IP 177/ASTM D664 method.

There are several strong arguments against the use of the IP 177/ASTM D664 method for Base Number: • The hydrochloric acid has an acid strength in the solvents used in this method, which only reacts with, and therefore determines “strong alkalinity,” >pH 11, in the lubricant sample. It does not determine “mild alkalinity,” up to pH 11, although it is not clear whether this is a crucial difference. • The method has poor reproducibility, although this is improved by using the replacement ASTM D4739 method, which uses a very slow potentiometric titration, 15 min/1 ml acid reagent added — an extremely slow method. • The sensitivity and fragility of the electrodes is important, the glass electrode is particularly fragile. Replacement glass electrodes must always be available, “conditioned” in the reaction solvent and ready for use. Another problem is that the electrode surfaces are gradually fouled by carbonaceous particulate in degraded lubricant samples and the electrode must be replaced. • This method is not unique as against the others, but all Base Number methods use chemicals with various forms of hazards, which are expensive to dispose of. The formal method uses a large test sample, 20 g, in 120 cm3 of solvent, the volume of which is increased by the ensuing titration. The test results are presented as milligrams of potassium hydroxide per gram sample equivalent. When applied to analyze successively degraded lubricant samples from engine bench or field tests, the IP 177/ASTM D664 Base Number method results tend to decline quickly in the initial stages of the test and then to decline more slowly, Figure 29.9, in contrast to results from other methods. It is generally held that a lubricant with a Base Number approaching a value of 2 should be replaced. Therefore, a Base Number of 2 for a degraded sample shows that its alkaline reserve equates to 2 mg potassium hydroxide per gram of sample. While the titration uses hydrochloric acid, this is related to its equivalent as potassium hydroxide. To sum up, the IP 177/ASTM D664 method suffers from the following: • • • •

Poor reproducibility, particularly for heavily degraded samples Lack of clarity in what it means Fragile apparatus Requiring large sample masses and solvent volumes

29.3.4.5 IP 276/ASTM D2896 Base Number By Perchloric Acid Titration This method is really a modification of the previous IP 177/ASTM D664 method, arising from the perception that changing the titrating acid from hydrochloric to the stronger perchloric, HClO4 , will react

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Handbook of Lubrication and Tribology

Base number

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IP276/D2896

IP 177/D664

Degraded lubricant, service life, h FIGURE 29.9 Base Number degradation values for same successive lubricant samples by IP 177/D664 and IP 276/D2896.

with both strong and mild alkalinity in lubricant samples. It is argued that the results using perchloric acid reflect the total additive content of the formulation. To accommodate the change in acid, the solvent must be modified as well and is a mixture of chlorobenzene and glacial acetic acid. The detection method is the same as for IP 177/ASTM D664, a combination of the glass electrode and the standard calomel electrode. The titration is the same and the plot of mV against acid volume has the same sigmoidal shape as given in Figure 29.7. Unfortunately, the method suffers from the same problems for heavily degraded samples, the plot then becoming indistinct with no clear endpoint, as in Figure 29.8. In this case, reproducibility is as poor as for the IP 177/ASTM D664 method. In this case, the method suggests a “back titration” with a much poorer range of reproducibility and repeatability. When this method is used to analyze degraded lubricant samples from engine bench or field tests, the IP 276/ASTM D2896 method Base Number results decline slowly throughout the test, in contrast to the results for the same samples analyzed using the IP 177/ASTM D664, as set out in Figure 29.6. There is no sharp decline in the initial stages of the test. There is a clear difference in results from the same samples between the IP 276/ASTM D2896 and the IP 177/ASTM D664 methods. As before, it is generally held that a lubricant with a Base Number approaching a value of 2 should be replaced. The test results have the same values as for the IP 177/ASTM D664 method. The solvents and chemicals used in IP 276/ASTM D2896 are even more hazardous and difficult/expensive to dispose of, than those used in the preceding IP 177/ASTM D664 method. The following summarizes, the IP 276/ASTM D2896 method: • It is a modification of the previous IP 177/ASTM D664 • It gives generally higher Base Number values, said to reflect the total, strong, and mild together, alkalinity present in a lubricant formulation • It has the same problem of an indistinct endpoint for heavily used samples • The solvents and chemicals used are hazardous and difficult/expensive to dispose of

29.3.4.6 ASTM D974 — Base Number by Color Indicator This method is worth noting but is now relatively little used. The method is very similar to IP 177/ASTM D664 method but uses an naphtholbenzein indicator color change to determine the neutralization endpoint. The results are expressed in the same way as IP 177/D664.

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Conductivity meter

Solution

Platinum electrodes

FIGURE 29.10

Diagram and picture of conductimetric cell.

29.3.4.7 IP 400 — Base Number by Conductimetric Titration IP 400 is relatively recent (there is no equivalent ASTM method) and directly addresses the problems of the previous methods. Chemically, it is identical to IP 177/ASTM D664 but the crucial difference is that it uses a conductimetric detection method to follow the progress of the neutralization reaction. It measures the resistance, or its inverse, the conductivity, of a solution between two platinum plates rigidly held in a glass tube, shown both as a diagram and picture in Figure 29.10. The plates are typically 10 mm square, welded to platinum wires, which exit through the wall of the glass tube to external connections. The conductimetric probes are very robust and work as well when bright metal or when coated with carbonaceous particulate. The only problem with electrode contamination occurs when the carbonaceous particulate coats the wall of the glass probe containing the electrodes sufficient enough to cause an electrical short circuit. The conductimetric cell does not need to be a special model, excellent results can be obtained using standard cells as used in initial physical chemistry laboratory experiments. Special cells are only constructed for automated systems, which use small volumes of sample, solvent, and titrating acid solution. The conductivity of the solvent plus sample is low, of the order of 2 µS (microsiemens) and increases linearly as the titration proceeds, Figure 29.7. At the endpoint, the gradient of the linear plot changes sharply. The endpoint is determined by the intersection of two linear plots, as shown in Figure 29.11.

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Handbook of Lubrication and Tribology 16 Conductance, mS

14

Conductance, mS

12 10 8 6 4 2 Endpoint 1.97 0

FIGURE 29.11

0

0.5

1

1.5 2 2.5 Volume HCI added, ml

3

3.5

4

Conductivity vs. titration volume plot for IP 400.

The crucial point about IP 400 is that the quality of the endpoint does not change with the sample condition, either new, lightly used, or heavily used, as shown in Figure 29.11. The intersection of the linear sections moves to lower values as the alkaline reserves of the samples reduce. Reproducibility is good, within the limits set for IP 276/ASTM D2896, with no deterioration with sample use, as shown. The test results are very close to those obtained using the IP 177/ASTM D664 method under its best conditions, which is not surprising as the chemistry is the same. Smaller volumes are used, the sample weight specified by the IP 400 procedure is 5 g but very good reproducibility and repeatability have been achieved down to 0.1 g for small volume and unique samples. The titration is relatively quick compared to the previous potentiometric methods, for during the titration the solution conductivity stabilizes as soon as the added aliquot of acid is thoroughly mixed. The IP 400 procedure is simple and straightforward. This demonstrates that the problems of the previous Base Number methods are associated with the following: • The potentiometric electrodes and their physical reactions. • The potential difference titrations of a number of substances in the used lubricant samples against their total conductivity. 29.3.4.8 Precision of Base Number Determinations The precision of these determinations has the following two forms, from cooperative test programs carried out between participating laboratories: 1. Repeatability, by the same operator, same laboratory 2. Reproducibility, by different operators and laboratories The format of precision is interestingly different from that normally encountered. It is set as a requirement that the results on the same sample should not vary by more than the stated limit values more than 19 cases out of 20, an interesting approach to a 95% confidence limit.

Repeatability For IP 177/ASTM D664, Base Number, By Manual Methods 7 mg Automatic Methods 6 mg

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Reproducibility 20 mg 28 mg

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Repeatability For IP 276/ASTM D2896, Base Number, New Lubricants 3% Used Lubricants 24% For IP 400 Base Number New and Used Lubricants 0.17x1/2 where x is the average of the results

29-27

Reproducibility 7% 32% with back titration 0.31x1/2

Note that there is no distinction in precision between new and used lubricant samples for the IP 400 conductimetric method for Base Number determination. 29.3.4.9 Fourier Transform Infrared Spectroscopy Methods Fourier transform infrared spectroscopy (FTIR), has been applied to the analysis of degraded lubricant and hydraulic fluids. The method is not direct in the sense of reading a value from a scale. The analysis is conducted indirectly by first obtaining various parameters derived from the difference spectra between the time sample and the original lubricant. Multivariate analysis and principal component regression (PCR), are then applied to these parameters to determine the Base Number. The overall process is now well established as a technique for measuring used lubricant properties. It has considerable potential as a nonwet, relatively “dry” method, which does not need to use hazardous laboratory chemicals. 29.3.4.10 Summary for Base Number Measurements For the reasons developed above, the “Base Number” value for a degraded lubricant is not a straightforward measurement. Any quoted values are not absolute and must be related to the method used to determine that value. The problems of the IP 177/ASTM D2896 and IP 276/ASTM D2896 methods lie with: • Repeatability/reproducibility difficulties introduced by the potentiometric electrode reactions. • Also, the interpretative differences seen between “strong” and “weak” alkalinity. The ASTM D974 colorimetric method is rarely used. The IP400 method is much better for Base Method measurement because of its clarity of endpoint, which is sustained for new, somewhat degraded and heavily degraded lubricant samples. The FTIR difference analyses combined with chemometric statistical analyses can predict Base Number and show considerable promise with very substantial reductions in the use of solvents and reactants. 29.3.4.11 Sources and Effects of Acidity In addition to acidity caused by combustion of inorganic compounds to give mineral acids, as described in Section 29.3.4.2, and also the oxidative degradation of hydrocarbon fuels and lubricants, hydraulic fluids will also degrade through localized high temperatures. Localized high thermal stress on a hydraulic fluid will, in due course but over a considerably longer period than for lubricants, cause thermal degradation and oxidation. These conditions will cause the physical properties of the hydraulic fluid to go outside its specification and it must be replaced. Acidity in degraded lubricant and hydraulic fluids corrodes system components. Corrosion combined with erosion gives enhanced wear rates, particularly in systems with mixed metals in contact by electrochemical effects. Corrosion can also generate solid debris within the system leading to clogging of tubes, filters, and obstruction of system operation. In collaboration with water, corrosion leads to rust formation.

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The areas that are prone to acidity attack are (i) bearing corrosion and (ii) cam and tappet corrosion and rusting. It may be surprising that acidity and alkalinity can exist together in a used lubricant. It is again helpful to go back to the explanation given in Section 29.3.4.2. The effect of hydrocarbons as solvents with their low permittivities (dielectric constants) has the effect of giving a greater range of acidity and also alkalinity to the components present in the system. Acid-base interactions can range from (i) complex formation, AHB- of acid, AH, and base, B-, together, with the acid proton shared between the acid and base, and (ii) to the full transfer of the proton from the acid to the alkali as normally understood by “neutralization.” In the former case, the substance can be both acid and basic (alkaline). Therefore, Base and Acid Numbers can coexist in the same system. Generally, in a new automotive lubricant sample, the Base Number will be high, of the order of 6–10 KOH units and the Acid Number can be of the order of 0.5–1.0 units. Acid Number is the corollary of Base Number but has not been subject to the same level of controversy as described for Base Number. Only recently, has there been difficulty with the Acid Numbers of synthetic ester lubricants used in gas turbine engines, addressed by an sampling and analytical error (SAE) method. Developments in applying conductimetric methods to the determination of Acid Number are also discussed as part of a method to determine Base and Acid Numbers sequentially for the same sample in the same apparatus.

29.3.4.12 Acid Number Determination by IP 177/ASTM D664 This method is directly analogous to the Base Number determination described previously in Section 29.3.4.4. The solvents for the sample are the same, a mixture of toluene, isopropyl alcohol, and water with the titrant being potassium hydroxide in alcohol. The method follows the neutralization of the sample solution by alcoholic alkali by using the glass and standard calomel electrode pair, giving a millivolt potential difference, between the electrodes against titration volume, V . The form of the titration is again a sigmoidal curve, with the endpoint at the change of gradient, the point of inflection, at the center of the sigmoidal plot. The endpoint is more clearly shown by the first derivative, d(mV )/dV plotted against V , Figure 29.12, which is an appropriate repeat of Figure 29.7.

Endpoint by first derivative

Electrode potential, mV

E(V )

E(mV )

E(V ) V(ml) Titration volume, ml

FIGURE 29.12 Acid Number determination by the IP 177/ASTM D664 method, mV vs. volume plot and first derivative plot.

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The test results are presented as milligrams of potassium hydroxide per gram sample equivalent, the same as for Base Number. The limits of repeatability and reproducibility are the same as for the determination of Base Number, as can be seen in the following table, as a requirement that the results on the same sample should not vary by more than the stated limit values more than 19 cases out of 20.

Repeatability By Manual Methods Automatic Methods

7 mg 6 mg

Reproducibility 20 mg 28 mg

29.3.4.13 D974 — Acid Number by Color Indicator Again, it is worth noting this method but it is relatively rarely used. It is very similar to IP 177/ASTM D664 but instead of a potentiometric method, it uses the color change of an indicator, naphtholbenzein, to determine the neutralization endpoint. The results are expressed in the same way, in milligram KOH per gram of sample. 29.3.4.14 Sampling and Analytical Error Determination of Acid Number in (Gas Turbine) Synthetic Ester-Based Lubricant Gas turbine lubricants are subjected to high temperatures in the center of the engine. But these lubricants are not exposed directly to combustion gases, as in a reciprocating engine. The high temperature within the central engine bearings causes breakdown of the esters to give acids and polyhydric alcohols and their degradation products. The organic acids give acidity to the ester lubricants, which cause corrosion to the engine unless controlled. It is important that this acidity is controlled for current gas turbines as future engines will operate at even higher internal temperatures and cause even more acidity. The results for Acid Number of the SAE method can be addressed by scrupulous attention to experimental detail in the IP 177/ASTM D664. 29.3.4.15 Simultaneous Conductimetric Determination of Base and Acid Numbers The conductimetric Base Number determination of IP 400 involves dissolving the lubricant sample in a toluene/isopropyl alcohol/water solvent and then titrating that solution with an alcoholic solution of hydrochloric acid, to give the well-known plot of Figure 29.13, Conductivity vs. Titration Volume. The sector A-B-C in Figure 29.13 is the Base Number titration, exactly the same as an IP 177/ASTM D664 titration for Base Number. This plot can be reversed by the addition of alcoholic alkali, which gives an almost exact symmetrically reversed plot, sector C-D-E. Further addition of alkali then titrates the original acidic content of the lubricant sample, sector E-F-G as the Acid Number titration. This method gives results within the IP and ASTM limits for repeatability. 29.3.4.16 Relationship Between Acid and Base Numbers of Degraded Lubricants The relationship between the Acid and Base Numbers for a degraded lubricant sample were developed for higher sulfur fuels and previous additive packages. Thus, the general rules, which developed were that if Acid Number rose to be greater than the declining Base Number, “crossing over,” then this was a condemning limit for the lubricant charge. Further, if Base Number declined below a value of 2, then this was a separate condemning limit for the lubricant charge. However, the gradual move to “low” and “lower” sulfur fuels for diesel fuel and, separately, modern additive packages can extend a system’s lubricant charge life. The condemning limits for degraded lubricants have changed considerably.

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Handbook of Lubrication and Tribology Br-tan oil Wayne kerr measurement of IP 177 BN/AN titration conductance term using isopropanol 50%/toluene 50% solvent

Mass of sample = 1.03 g BN

Ω–1 × 10–6

2.5

endpoint = 1.125 ml BN = 5.8

2.0 AN 1.5

endpoint = 2.65–2.12 = 0.53 ml AN = 2.8

1.0 0.5

1 V/ml HCI

2

1

2 V/ml KOH

3

4

Explanation of the BN/AN back titration curve KOH

Conductance

HCL

A

B

C Volume

D

E

F

The titration curve is seen to pass through a maximum of six different regions. The region up to point B represents the addition of HCI and from point B to point F represents the addition of KOH.

FIGURE 29.13

Sequential conductimetric acid and base titration of lubricant sample.

29.3.5 Water Content Water commonly contaminates machinery lubricant and hydraulic systems; its presence reduces the load carrying ability of a lubricant and increases wear. In addition, it promotes oxidation and corrosion. For synthetic polyol esters, water degrades the base stock back to its component acid and polyol. Maximum “safe” levels of water are usually taken to be 0.1–0.2%, higher for engines, lower levels for machinery and hydraulic systems. Water contamination of engine lubricant and hydraulic systems commonly arises from the following: • Combustion water, recalling that hydrocarbon combustion gives carbon dioxide and water as products. Some of the water passes into the crankcase as “blowby” down the side of the piston and condenses at the lower temperatures of that region of the engine. • Condensation of water in engines or hydraulic systems on standing or condensation into fuel tanks/hydraulic fluid reservoirs when operating at low/very low ambient temperatures. • Leakage into the fluids from cooling systems, such as circulating cooling water in engines by gasket failure, or leakage within the matrix of a heat exchanger. Almost all heat exchangers leak to some

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extent, acceptable ones leak very, very little but the leakage rate eventually increases with corrosion to become significant. • Water in lubricants degrades their formulation by absorbing acid gases to form strong acids. The presence of water in formulations can cause the additive package to precipitate out (“drop out”) as a severe form of degradation, which leaves the base oil only to lubricate the system. There are various methods to determine water in hydrocarbons and also lubricant and hydraulic fluids for the following reason: • The different nature of relatively pure hydrocarbons, as fuels, and lubricant and hydraulic fluids as complex formulations. • The different nature of the physical methods used to determine water in these fluids. • The varying nature of water at different concentrations. Water is a very complex physical substance for which complete models have yet to be accepted. Many models have been proposed for the physical properties of water but it is clear that “bulk” water, as a large polymeric but transient structure, has different physical properties from smaller groups of water molecules or, indeed, individual water molecules. Indeed, the infrared spectrum of very dilute water in organic solvents is a relatively narrow band centered on one frequency, whereas higher concentrations have appreciably wider bands at a shifted frequency. From these considerations, the various viewpoints of the methods used to measure the water content of lubricant and hydraulic fluids can be appreciated. 29.3.5.1 The IP 74/ASTM D95 — Water in Petroleum Products and Bituminous Materials by Distillation (the “Dean and Stark” method) The IP 74/ASTM D95 (“Dean and Stark Method”) for the determination of water in hydrocarbon fluids is a “total” method, rather gross and sensitive up to the 12% level. The method selectively distils water from petroleum products to separate and measure it using an organic solvent. It is an applied steam reflux distillation, which separates and concentrates the condensed water into a separate, calibrated, test tube, Figure 29.14. One problem in measuring the volume of water is complete separation of the water and hydrocarbon in the calibration test tube, which can be clear (complete) or hazy (incomplete), dependent upon the nature of the fluids and additive components present. The glassware apparatus for the Dean and Stark distillation is shown in Figure 29.11, note that the calibrated test tube in the system is positioned such that the water evaporated from the hydrocarbon fluid sample is collected and measured. The Dean and Stark method can be seen as a “total” water determination method as it collects all of the water from the sample that can be volatilized. Its limitation is that it uses an equilibrium water distribution between the (sample + organic solvent) and the (water + organic solvent), thus almost all of the water is removed to the measurement calibrated test tube. 100 g of oil sample is continuously distilled/refluxed with ∼100 cm3 of xylene, an aromatic solvent immiscible with water. The procedure is continued for 1.5 to 2 h to ensure that all of the water has been transferred. The percentage of water present in the sample is expressed as the volume of water in the graduated test tube multiplied by 100% and divided by the mass of oil sample. The method is direct with an unequivocal measurement of water but has the following disadvantages: • Lack of sensitivity • Occasional problems of measuring the water content because of incomplete separation of water/xylene in the measuring tube • The time of measurement, upward of 1.5 to 2 h per sample • Personnel intensive

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Condenser

Receiver

Distillation vessel

FIGURE 29.14

The Dean and Stark method apparatus for the determination of water in hydrocarbon fluids.

29.3.5.2 IP 356/ASTM D1744 — Determination by Karl Fischer Titration The Karl Fischer method of water determination is frequently discussed and results from it often quoted. It uses the reaction of water with iodine and sulfur dioxide in a pyridine/methanol solution, which is unpleasant to use. Iodine in a methanol/chloroform solution is an alternative reagent. The reagent reacts with hydroxyl groups, –OH, mainly in water but also in other hydroxylic compounds such as glycol, CH2 OH–CH2 OH, and depolarizes an electrode. The resulting potentiometric change is used to determine the endpoint of the titration and thus calculate the concentration of water in the oil sample. While the Karl Fischer method might be used to determine the water content of a formulated lubricant or hydraulic sample, it has never been approved for this purpose. The method was originally developed to determine the concentration of water in crude oil and can be used to determine water in fuels. When used to determine water in new and degraded formulated lubricants and hydraulic fluids, the method overdetermines a “water response” because the reagent not only reacts with water but also some of the additives present. This is a problem, because the Karl Fischer response for a new oil can give a blank value of 2%, mainly from the additive pack. But a failure limit for water in internal combustion engines is typically set at 0.2% or lower, thus the failure limit is an order of magnitude less than the blank value. Worse, however, is the problem of the oil additives degrading during service life, which may form unknown compounds which may or may not react with the Karl Fischer reagent. The “blank value” is now in doubt for used samples, it can be estimated but this leaves a possibly large margin of error. Therefore, the Karl Fischer titration method for the determination of water in new and degraded formulated lubricant and hydraulic fluids is fraught with difficulty. Some variations have been tried, such as gently sparging the oil sample with dry nitrogen and thus blowing the water content as vapor over into a Karl Fischer titration. This takes a long time to “complete” and it is uncertain at what point, if complete at all, when “all” of the water has been transferred for measurement. The Karl Fischer determination of water in formulated new/degraded lubricant and hydraulic fluids is unsuitable because of reactions with additives. When additives are absent, then the Karl Fischer method is a sound method to determine the water content of “pure” hydrocarbon fluids, such as base oils.

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The Degradation of Lubricants in Service Use

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Bio-Rad Win-IR 0.4% 0.3 Absorbance

0.3% 0.2

0.2% 0.1%

0.1

0.05% 0.0%

0 3800

3600

3400

3200

Wavenumber (cm–1)

FIGURE 29.15 FTIR spectra of hydrocarbon fluids degraded with water. From Machinery Oil Analysis — Methods, Automation and Benefits, 2nd ed., Larry A. Toms, Coastal Skills Training, Virginia Beach, Virginia, USA (1998). Courtesy Bio-Rad Laboratories.

29.3.5.3 Water Content by FTIR Spectrophotometry The O–H group in water has a strong, broad, and distinctive infrared absorption from 3150 to 3500 cm−1 , centered on 3400 cm−1 . The absorption band is broad because the O–H group is hydrogen bonded for groups of water molecules. As the concentration of water decreases it becomes less hydrogen bonded or exists as more as smaller groups of bonded molecules or even individual molecules, and its molar absorption increases. Therefore, the calibration curve tends to be nonlinear at lower concentrations. A representative set of FTIR spectra for different levels of water contamination of hydrocarbon oils is given in Figure 29.15 and inspection demonstrates the nonlinearity of the water absorbance in the 0.0 to 0.2% concentration range. The method works well except for formulations using polyol ester base oils or with high dispersant/detergent additive levels. In the first case, the problem arises from the polyol ester infrared absorption in the previously used 3150–3500 cm−1 region, centered on 3400 cm−1 . Subject to detailed baseline corrections, the 3595 to 3700 cm−1 region is used instead to determine water contamination in these oil samples, this corresponds to a singly bonded O–H group. For high detergent/dispersant lubricant samples, the hydroxyl absorption band is not seen but a background increase in absorption occurs between 3000 and 4000 cm−1 . This effect is nonlinear and must be calibrated with standard solutions. It is separate from baseline shifts due to soot and particulates, which are unlikely to be present in this type of lubricant formulation. These two effects point to the main limitation of the FTIR method, which essentially reduces to the need to know the nature of the fresh, unused lubricant. This means that the FTIR method cannot be applied universally and will give errors occasionally, when samples of oils based on polyol esters or formulations containing high levels of dispersants/detergents. Other than this limitation, the FTIR method is very useful and shows great potential for the rapid and accurate determination of water degradation in many new and used lubricant formulations.

29.4 Minor Methods of Investigating Lubricant Degradation Description of the following methods as“minor”means that they are only used in particular and individual circumstances to investigate the degradation of lubricants and hydraulic fluids. They do not form part of a routine investigation of degraded lubricant samples.

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29.4.1 Density, “Gravity,” or “Specific Gravity” Density of a lubricant sample is also referred to as its “Specific, or API, Gravity” and has little value as a measure of the degraded lubricant’s fitness for purpose. The determination of lubricant sample density is now be readily measured to three significant figures using vibrating tube detectors, a very much shorter and accurate procedure than the density bottle method or glass hydrometer, ASTM D1298. But the density information gained has little importance, the density of a degraded lubricant should be close to that of the original material. Changes in density show contamination by a solvent, such as fuel dilution, a different product inadvertently added or a build-up of foreign material. The differences are nevertheless small and, as an example, fuel dilution needs to be extensive to see a significant change in density.

29.4.2 Flash Point of Degraded Lubricant Flash point determination of lubricant samples can now be considered more readily due to automated instruments now being readily available. Both manual and automated methods are based upon the PenskyMartins method, as in ASTM D93 for diesel lubricants. The method brings together considerations of volatility, combustion limits, and ignition temperatures to give a useful measure of great utility. Flash point values of degraded samples rarely increase, if they do, a higher viscosity fluid has been inadvertently added. Much more likely is a decrease in flash point for a degraded lubricant sample caused primarily by fuel dilution resulting from cold/low temperature engine operation. Thermal decomposition of the base oil under extended power operation may also generate lighter fractions which reduce the sample flashpoint. Reduced flashpoints of degraded diesel lubricants due to fuel dilution would normally be associated with a decreased viscosity value and a crosscheck should be done for this. The quantitative extent of fuel dilution is usually non-linear with respect to flashpoint and should be measured by either gas chromatogrphy or FTIR methods. Various method procedures exist of increasing accuracy, the Cleveland open cup, ASTM D92, the Pensky-Martins closed cup, ASTM D93 and the Setaflash small scale closed cup, D3828. Flash points for degraded petrol/gasoline fuel dilution in degraded lubricants are measured by ASTM D322.

29.4.3 Foaming of Lubricants Lubricant foam has a low load carrying ability. Excessive foam build-up in a reservoir or sump will rapidly lead to excessive wear and catastrophic failure of the system. Too high a level of lubricant in an engine sump, by overfilling or miscalibration of the level indicator (dipstick) causes the crankshaft and connecting rod big-end caps to whip up the lubricant into an all-pervading foam and rapid damage ensues. Air leaks into the oil flow or an open drop from a supply pipe into a hydraulic fluid reservoir can generate foam. Operationally, engines should not be overfilled, the level indicator correct, leaks stopped, and supply pipes extended to deliver return lubricant below the normal liquid surface level in a reservoir. While base oils have little foaming tendency, modern lubricant formulations contain many additives substances such as detergents, which can enhance their tendency to foam. Surface active additives will also increase the foaming tendency of a formulation. ASTM D892 measures the foaming tendency of a hydrocarbon fluid but is much more relevant to the fresh, unused, material under laboratory conditions than degraded samples in operating systems. Resolving a used lubricant foaming problem should be treated with great care, fortunately it is relatively rare. Foaming of the new formulation is controlled by the addition of liquid silicone polymers, which reduces the surface tension at the contact points of the foam cells. This allows the lubricant to drain away and the foam to subside. However, formulating the optimum silicone concentration requires extensive work as too little or too much silicone additive increases the foaming tendency of the formulation. Adding an antifoam silicone liquid in situ to a foaming, degraded lubricant, or hydraulic fluid should be approached very carefully and incrementally.

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29.4.4 System Corrosion (“Rusting”) with Degraded Lubricants Lubricant formulations contain rust inhibitors and a system which is maintained well, with a maximum water content of 0.1% for engines and 0.01% for other systems, should not have rust problems. Corrosion inhibitors are needed for systems and vehicles that are used very intermittently, such as military vehicles in prolonged storage or vehicles delivered from one side to the world to another. Repeated sequences of cold engine starts and very short drive distances up to dealer delivery causes condensation of water in the engine and ensuing corrosion of ferrous parts. The contaminants building up in a degraded lubricant system can adversely affect the action of the rust inhibitor present by competitive adsorption at metal surfaces. The ASTM D665 rust method applies to lubricant formulations, if rust corrosion is either suspected or present by discoloration, then degraded samples should be sent for test.

29.4.5 Demulsibility and Interfacial Tension of Degraded Lubricants The demulsibility characteristic of a lubricant is its ability to separate from water when emulsions are formed in a system. While the test is performed for new lubricants, the build-up of trace contaminants may reduce the separation from water in emulsions for degraded samples, hence the term “demulsibility.” Testing degraded lubricants for demulsibility in the laboratory may not be indicative of that sample’s performance in operating systems. ASTM D1401 is the demulsibility test for turbine lubricants and ASTM D2711 is for medium- and high viscosity lubricants. A slightly different procedure of ASTM D2711 is used for extreme pressure (EP) lubricants. Interfacial tension (IFT), is a measure of the surface energy of a fluid against a solid surface or an immiscible standard fluid. Additives contribute to that surface energy and a decrease in interfacial tension indicates that these additives are being deactivated, removed in some way, or depleted by oxidation. A decrease in interfacial tension is an early indication of oxidation before changes are noticed in Acid Number or viscosity. Alternatively, the circulating lubricant is collecting certain compounds in the system added as rust inhibitors, which have polar structures. Interfacial tension measurements of degraded lubricants are useful for rust- and oxidation-inhibited turbine and transformer oils by ASTM D971. The results should not be interpreted on their own but related and compared to changes in other measurements of the system, particularly viscosity and Acid Number.

29.4.6 Instrumental Analytical Techniques The spectroscopic, chromatographic, and x-ray analytical techniques represented by FTIR, gas and liquid chromatography (GC/LC), and x-ray diffraction (XRD/XRF) are increasingly used to investigate degraded lubricants. The long-term trend is for the cost of the instruments to decrease and their resolution to increase with enhanced information technologies. Fourier transform infrared technique is increasingly used to analyze degraded samples, particularly for sequential samples compared to new, unused, samples of the same lubricant. Selected regions of the infrared region are used to follow particular aspects of sample degradation. The method is given additional power through the use of multivariate data analysis. Toms describes the application of FTIR to the analysis of degraded samples. Chromatography, particularly liquid chromatography, may be used to analyse additives in lubricant formulations. CEC has very high resolution of additives and can follow their depletion in successive degraded samples. Gel permeation chromatography (GPC), can follow the degradation or scission, of polymer chain lengths and therefore mean molecular weights, of additives such as VIIs, dispersants, and other polymeric additives. X-ray diffraction is mainly used for quality control and to identify unknown deposits; of more importance is XRF and x-ray absorption fine structure (XAFS), used to identify the elements in compounds, liquids, and solids, found in operating systems.

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These instrumental techniques will have increasing importance in the analytical investigation of degraded lubricant samples. An important issue, already begun, is to bring these instrumental analytical techniques and their specific application developments into the set of standard analytical method for the lubricant and hydraulic fluid manufacture and service use industries.

29.5 Case Studies of Degraded Lubricants 29.5.1 A Degraded Lubricant Sample from a Heavy Duty Diesel Engine

New

Used

Standard

Odor

mid-brown, transparent mild

black, ← opaque diesel, slightly burnt ←

Viscosity at 40◦ C, mm2 /sec at 100◦ C Viscosity index Acid Number, mg/g/KOH Base Number, mg/g/KOH Water Percentage of Soot Percentage of Fuel Dilution

71.31 11.71 160 2.8 9.6 nil 0 0

61.82 10.58 162 4.5 4.5 nil 1.2% 4%

ICP Elements, mg/kg or ppm P Zn Ca Ba B Mg Na Fe Al Cr Cu Pb Sn Ni Mo Si

350 400 1100