The Handbook of Tribology Data: An excellent Friction, Lubrication and Wear Resource [1 ed.] 0849339049, 9780849339042, 1420050478, 9781420050479

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TRIBOLOGY DATA HANDBOOK EDITED BY

E. Richard Booser An Excellent Friction, Lubrication and Wear Resource SPONSORED BY THE

Society of Tribologists and

Lubrication Engineers

Copyright © 1997 CRC Press, LLC.

Library of Congress Cataloging-in-Publication Data Tribology data handbook / edited by E. Richard Booser. p. cm. Includes bibliographical references and index. ISBN 0-8493-3904-9 (alk. paper) 1. Tribology—Handbooks, manuals, etc. I. Booser, E. Richard. TJ1075.T755 1997 621.8’0—dc21 97-6215 CIP 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 authors and the publisher cannot assume responsibility for the validityof all materials or for the consequences of their use. Neither this book nor any part may be reproduced or transmitted in any form or by any means, electronic or mechanical, including photocopying, microfilming, and recording, or by any information storage or retrieval system, without prior permission in writing from the publisher. All rights reserved. Authorization to photocopy items for internal or personal use, or the personal or internal use of specific clients, may be granted by CRC Press LLC, provided that $.50 per page photocopied is paid directly to Copyright Clearance Center, 222 Rosewood Drive, Danvers, MA 01923 USA The fee code for users of the Transactional Reporting Service is ISBN 0-8493-3904-9/97/$0.00+$.50. The fee is subject to change without notice. For organizations that have been granted a photocopy license by the CCC, a separate system of payment has been arranged. The consent of CRC Press LLC does not extend to copying for general distribution, for promotion, for creating new works, or for resale. Specific permission must be obtained in writing from CRC Press LLC for such copying. Direct all inquiries to CRC Press LLC, 2000 N.W. Corporate Blvd., Boca Raton, Florida 33431. Trademark Notice: Product or corporate names may be trademarks or registered trademarks, and are used only for identification and explanation, without intent to infringe.

Visit the CRC Press Web site at www.crcpress.com © 1997 by CRC Press LLC. No claim to original U.S. Government works International Standard Book Number 0-8493-3904-9 Library of Congress Card Number 97-6215 Printed in the United States of America 2 3 4 5 6 7 8 9 0 Printed on acid-free paper

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Preface As a source book on friction, lubrication and wear, this Handbook is sponsored by the Society of Tribologists and Lubrication Engineers (STLE) to consolidate information on tribology. For detailed discussions of specific topics, many users will likely continue to find it helpful to refer to these earlier volumes of this series: Vol. I. Handbook of Lubrication. Application and Maintenance. 1983. Vol. II. Handbook of Lubrication. Theory and Design. 1984. Vol. III. Handbook of Lubrication and Tribology. Monitoring, Materials, Synthetic Lubricants, and Applications. 1994. While the earlier volumes provided broad coverage, this present Handbook is intended to provide a single source of “look-up” information for the entire field of tribology in tables of data, design charts, equations, and recent references. Any related discussions are generally limited to essentials required to introduce or explain the information involved. In addition to the specialized background of the 77 contributors, several other sources of information warrant mention. Producers of mineral oil and synthetic lubricants, solid film lubricants, bearings, gears, and various machinery supplied typical data and specifications. Unique numeric compilations of lubricant properties, seal material behavior, and friction and wear data are included from projects sponsored by the National Institute of Science and Technology for a computerized tribology information system (ACTIS). Several portions of the Handbook are of special interest. An extensive set of current military lubricant specifications in Chapter 18 is followed by a variety of industrial and automotive specifications in the remainder of Sections II and III. Sections V, VI and VII reflect facets of friction and wear for which basic understandings are still evolving. Performance and design relations are collected in Section VIII for machine elements such as bearings, seals, gears, brakes, piston rings, and magnetic storage units for computer systems. Section IX provides similar data for metal working. Perhaps the most challenging coverage is given in the final two Sections. Section X gives comprehensive guidelines for monitoring, maintenance, and failure assessment in lubrication of automotive, industrial and aircraft equipment. Environmental and toxicology questions in Section XI appear to be of somewhat less current concern, even while under close scrutiny and control by the government. Developed with guidance and participation by the STLE Technical Committees and Industry Councils, it is hoped that this Tribology Data Handbook will provide a useful aid for those working to achieve more effective lubrication, better control of friction and wear, and improved understanding of the complex factors involved in tribology. E. R. Booser Editor

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The Editor Dr. E. Richard Booser has been active in work on tribology and lubrication for over 50 years. As an instructor and research assistant in chemical engineering at The Pennsylvania State University, his early research studies focused on refining procedures and performance testing for petroleum lubricants. He was then employed by the General Electric Co. for 39 years in development work on bearings and lubricants for steam and gas turbines, electric motors and generators, aerospace and nuclear plant equipment, and a variety of related electrical products. He currently works as a consulting engineer on bearings and lubrication. Assignments have covered lubrication for nuclear power plants; bearing performance and problem analyses for turbines, generators, electric motors and accessory power plant equipment; friction and wear testing; locomotive, aerospace, and appliance bearings; and failure analysis of ball and roller bearings. His 85 publications cover properties and performance of lubricants and bearing materials, turbulence and parasitic power loss in high-speed oil-film bearings, oil oxidation, grease life in ball bearings, circulating oil systems, and lubrication of electric motors. In addition to his work as editor of the earlier three volumes in this CRC Press handbook series, he also co-authored the 1957 McGraw-Hill book on Bearing Design and Application. While President of the Society of Tribologists and Lubrication Engineers in 1956, he participated in initiating the annual joint Tribology Conferences with ASME. He has also organized and taught bearing and lubrication courses and seminars for over 800 designers and engineers. In 1992 he received the STLE National Award. Dr. Booser and those contributing to the Handbook have drawn on their own experiences and on the resources of STLE to develop this compilation of information in the emerging field of tribology: the science of friction, wear, and lubrication.

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Table of Contents Section I. Lubricant Properties 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16.

Typical Lubricating Oil Properties.................................................................................3 Chun-I Chen Comparison of Properties of Synthetic Fluids..........................................................34 Wilfried J. Bartz Polyalphaolefins...............................................................................................................37 Ronald L. Shubkin Organic Esters.................................................................................................................42 Bruce J. Beimesch Polyalkylene Glycols.......................................................................................................49 William L. Brown Phosphate Esters.............................................................................................................60 Douglas G. Placek and M. P. Marino Silicones............................................................................................................................75 E. D. Brown Perfluoroalkylpolyethers.................................................................................................80 Gregory A. Bell Polychlorotrifluoroethylenes..........................................................................................88 Louis L. Ferstandig Polybutenes......................................................................................................................94 John D. Fotheringham Vegetable Oils — Structure and Performance.........................................................103 Saurabh S. Lawate, Kasturi Lal, and Chor Huang Additives — Chemistry and Testing..........................................................................117 Syed Q. A. Rizvi Greases............................................................................................................................138 Richard C. Schrama Solid Lubricants.............................................................................................................156 E. Richard Booser Viscosity of Liquids, Aqueous Solutions and Liquid Metals.................................159 (CRC Press) Gas Properties...............................................................................................................169 Donald F. Wilcock

Section II. Typical Lubrication Specifications for Equipment 17. 18. 19. 20.

Comparison of Viscosity Classifications..................................................................179 E. Richard Booser Military Specifications..................................................................................................182 Bobby D. McConnell Air Compressor Lubrication.......................................................................................242 William Scales Automatic Transmission Fluids..................................................................................248 (Lubrizol Corp.)

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21. 22. 23. 24. 25. 26. 27.

Automotive Engine Lubricants........................................................................253 Yeau-Ren Jeng and Ron Desing Diesel Engine Lubricants..................................................................................263 Ron Desing Electric Motors and Generators......................................................................272 E. Richard Booser Hydraulic and Turbine Oils..............................................................................274 Ian Macpherson and Daniel D. McCoy Marine Diesel Engines......................................................................................282 (Lubrizol Corp.) Railroad Diesel Engines....................................................................................284 (Lubrizol Corp.) Slideways..............................................................................................................286 Ian Macpherson

Section III. Industrial Application Practices 28. 29. 30. 31. 32. 33. 34. 35. 36.

Aluminum Production Equipment.................................................................289 Edward J. Myers Farm and Off-Highway Machinery.................................................................298 (Lubrizol Corp.) Food Industry Lubrication...............................................................................303 George R. Arbocus Forest and Paper Products...............................................................................309 Frederick J. Villforth, III Machine Tools....................................................................................................320 Douglas Vallance Mining Excavation Equipment Lubricant Guidelines.................................332 Charles D. Barrett Refrigeration and Air Conditioning Lubricants............................................342 Kenneth C. Lilje, Thomas E. Rajewski, and Edan E. Burton Space and High Vacuum Lubricants...............................................................355 Robert L. Thorn and Michael R. Hilton Steel Industry Lubricant Guides......................................................................364 Richard C. Schrama and Daniel D. McCoy

Section IV. Lubricant Application Systems 37. 38. 39.

Centralized Lubrication for Industrial Machines..........................................385 James H. Simpson, III Oil Mist................................................................................................................396 Stanley C. Reiber Circulating Oil Systems.....................................................................................404 E. Richard Booser

Section V. Friction, Wear, and Surface Characterization 40. 41.

Surface Texture...................................................................................................415 V. V. Dunaevsky, Y.-R. Jeng, and J. A. Rudzitis Typical Friction and Wear Data.......................................................................435 A. W. Ruff

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42. 43. 44. 45. 46.

Friction and Wear Equations.........................................................................................445 Valery V. Dunaevsky Generalized Wear Coefficients......................................................................................455 Valery V. Dunaevsky Friction Temperatures.....................................................................................................462 Valery V. Dunaevsky Boundary Lubrication Relations....................................................................................474 Richard S. Fein Lubricated Wear Problems — Symptoms and Prevention.......................................486 Douglas Godfrey

Section VI. Material Properties 47. 48. 49. 50. 51. 52. 53. 54.

Typical Properties of Sliding Contact Materials.........................................................493 A. W. Ruff Rolling Element Bearing Materials...............................................................................495 Charles A. Moyer Oil Film Bearing Materials.............................................................................................503 George R. Kingsbury Mechanical Properties of Gear Materials....................................................................526 Lewis Rosado Wear-Resistant Hard Materials.......................................................................................540 William A. Glaeser Friction, Wear and PV Limits of Polymers and Their Composites........................547 Thierry A. Blanchet Properties of Advanced Ceramics................................................................................563 S. Frank Murray Friction and Wear Tests on Advanced Ceramics........................................................573 Amp Gangopadhyay

Section VII. Tribological Surface Modifications 55. 56. 57.

Tribological Surface Treatments and Coatings...........................................................581 Francis E. Kennedy and Ursula J. Gibson Friction and Wear of Hard, Thin Coatings.................................................................594 Arup Gangopadhyay Bonded Solid Film Lubricants.......................................................................................600 Robert M. Gresham

Section VIII. Component Performance and Design Data 58. 59. 60. 61. 62.

Fundamentals of Elastohydrodynamic Lubrication..................................................611 Michael M. Khonsari and D. Y. Hua High Pressure Viscosity and EHL Pressure-Viscosity Coefficients........................638 Richard S. Fein Rolling Bearings Performance and Design Data........................................................645 Thomas Jendzurski and Charles A. Moyer Journal Bearing Design and Analysis...........................................................................669 Michael M. Khonsari Design and Analysis of Hydrodynamic Slider Thrust Bearings..............................680 Michael M. Khonsari and J. Y. Jang

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63. 64. 65. 66. 67. 68. 69. 70.

Squeeze-Film Bearings....................................................................................................691 M. M. Khonsari and J. Y. Jang Ring- and Wick-Oiled Starved Journal Bearings........................................................708 Richard C. Elwell Porous Metal Bearings....................................................................................................719 Cris Cusano Dynamic Seals..................................................................................................................734 Alan O. Lebeck, for the Seals Technical Committee, STLE Gear Lubricant Selection and Application..................................................................781 Robert L. Errichello Brakes.................................................................................................................................800 Valery V. Dunaevsky Piston Rings......................................................................................................................811 Valery V. Dunaevsky Magnetic Storage System Materials...............................................................................823 Zunde Yang and Yip-Wah Chung

Section IX. Metalworking 71. 72. 73. 74.

Metal Removal: Fluid Selection and Application Guide...........................................831 Gregory J. Foltz and Harold J. Noble Metal Forming..................................................................................................................840 Ted G. McClure Microbial Control of Coolants......................................................................................854 Alan C. Eachus Additives for Metalworking Fluids................................................................................862 Neil M. Canter

Section X. Monitoring, Maintenance, Failure Patterns 75. 76. 77. 78. 79. 80. 81. 82. 83. 84.

Elements of an Oil Analysis Program.........................................................................875 James C. Fitch Wear Monitoring and Metal Analysis...........................................................................889 Costandy S. Saba Laboratory Used Oil Analysis Methods.......................................................................897 Mate Lukas and Daniel P. Anderson Gas Turbine Engine Lubricant Monitoring and Analysis.........................................915 Costandy S. Saba Automotive Engine Oil Life Factors............................................................................923 Shirley E. Schwartz Automotive Engine Oil Condition Monitoring..........................................................927 Donald J. Smolenski and Shirley E. Schwartz Diesel Engine Used Oil Analysis..................................................................................935 Jack Poley Rotating Machinery Vibration and Condition Monitoring.......................................944 William D. Marscher Ball and Roller Bearing Troubleshooting.....................................................................956 Thomas Jendzurski Gear Distress and Failure Modes..................................................................................986 (Falk Corp.)

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85. 86.

Filtration and Particle Count Classifications..........................................................1009 Charles A. Moyer Life of Oils and Greases...........................................................................................1018 E. Richard Booser

Section XI. Toxicology, Environment, Safety and Health 87.

Toxicology, Environment, Safety and Health........................................................1031 Joseph M. Perez and Donald I. Hoke

Appendixes 1. 2. 3.

Material Hardness Tables, Tests, and Data.............................................................1049 Charles A. Moyer Viscosity Conversion Factors....................................................................................1063 Douglas Godfrey International System of Units (SI) and Conversion Factors...............................1069 (CRC Press)

Index..........................................................................................................................................1087

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Advisory Board George R. Arbocus Consultant Elf Lubricants North America Keystone Division Linden, New Jersey Michael M. Khonsari, Ph.D. Chairman Department of Mechanical Engineering Southern Illinois University Carbondale, Illinois Daniel D. McCoy Industrial Products The Lubrizol Corporation Wickliffe, Ohio

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Charles A. Moyer (Retired) The Timken Company Canton, Ohio Joseph M. Perez, Ph.D. Professor of Tribology Department of Chemical Engineering The Pennsylvania State University University Park, Pennsylvania Ronald L. Shubkin, Ph.D. Research and Development Albemarle Corporation Baton Rouge, Louisiana

Contributors Daniel P. Anderson Spectro Inc. Littleton, Massachusetts

Neil M. Canter Chemical Solutions Willow Grove, Pennsylvania

George R. Arbocus Elf Lubricants North America Linden, New Jersey

Chun-I Chen National Institute of Science and Technology Gaithersburg, Maryland

Charles D. Barrett Castrol Industrial North America Downers Grove, Illinois Wilfried J. Bartz Technische Akademie Esslingen Osfildern Germany Bruce J. Beimesch Henkel Corp. — Emery Chemicals Cincinnati, Ohio Gregory A. Bell Dupont Technical Laborataory Chambers Works Deepwater, New Jersey Thierry A. Blanchet Dept. of Mechanical Engineering Rensselaer Polytechnic Institute Troy, New York E. Richard Booser Scotia, New York E. D. Brown Schenectady, New York William L. Brown Union Carbide Corporation Tarrytown, New York Edan E. Burton CPI Engineering Services Midland, Michigan

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Yip-Wah Chung Dept. of Materials Science and Engineering Northwestern University Evanston, Illinois Cris Cusano Dept. of Mechanical Engineering University of Illinois Urbana, Illinois Ron Desing The Lubrizol Corporation Wickliffe, Ohio Valery V. Dunaevsky Allied Signal, Inc. Elyria, Ohio Alan C. Eachus Angus Chemical Company Buffalo Grove, Illinois Richard C. Elwell Niskayuna, New York Robert L. Errichello Geartech Townsend, Montana Richard S. Fein Fein Associates Poughkeepsie, New York Louis L. Ferstandig Halocarbon Products Corp. River Edge, New Jersey

James C. Fitch Diagnetics, Inc. Tulsa, Oklahoma

Thomas Jendzurski SKF USA, Inc. King of Prussia, Pennsylvania

Gregory J. Foltz Cincinnati Milacron Cincinnati, Ohio

J. Y. Jang Dept. of Mechanical Engineering University of Pittsburgh Pittsburgh, Pennsylvania

John D. Fotheringham BP Chemicals Grangemouth Stirlingshire, Scotland Arup Gangopadhyay Ford Motor Company Dearborne, Michigan Ursula J. Gibson Thayer School of Engineering Dartmouth College Hanover, New Hampshire William A. Glaeser Battelle Columbus, Ohio Douglas Godfrey Wear Analysis San Rafael, California Robert M. Gresham E/M Corporation West Lafayette, Indiana Michael R. Hilton Aerospace Corp. Los Angeles, California Donald I. Hoke The Lubrizol Corporation Wickliffe, Ohio D. Y. Hua Center for Engineering Northwestern University Evanston, Illinois Chor Huang The Lubrizol Corporation Wickliffe, Ohio

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Yeau-Ren Jeng Dept. of Mechanical Engineering National Chung Cheng University Chia-Yi, Taiwan Francis E. Kennedy Thayer School of Engineering Dartmouth College Hanover, New Hampshire Michael M. Khonsari Dept. of Mechanical Engineering Southern Illinois University Carbondale, Illinois George R. Kingsbury Consulting Engineer Lyndhurst, Ohio Kasturi Lal The Lubrizol Corporation Wickliffe, Ohio Saurabh S. Lawate The Lubrizol Corporation Wickliffe, Ohio Alan O. Lebeck Mechanical Seal Technology Albuquerque, New Mexico Kenneth C. Lilje CPI Engineering Services Midland, Michigan Malte Lukas Spectro, Inc. Littleton, Massachusetts Ian Macpherson Ethyl Corporation Richmond, Virginia

Michael P. Marino FMC Corporation (Ret.) Philadelphia, Pennsylvania

Stanley C. Reiber Alemite Corp. Charlotte, North Carolina

William D. Marscher Mechanical Solutions, Inc. Parsippany, New Jersey

Syed Q. A. Rizvi The Lubrizol Corporation Wickliffe, Ohio

Ted G. McClure Fuchs Lubricants Company Harvey, Illinois

Lewis Rosado Aero Propulsion and Power Directorate Wright Laboratory Wright-Patterson AFB, Ohio

Bobby D. McConnell Bluff City, Tennessee Daniel D. McCoy The Lubrizol Corporation Wickliffe, Ohio

J. A. Rudzitis Dept. of Production Engineering Riga Technical University Latvia

Charles A. Mover The Timken Company (ret.) North Canton, Ohio

A. W. Ruff National Institute of Science and Technology (ret.) Gaithersburg, Maryland

S. Frank Murray Dept. of Mechanical Engineering Rensselaer Polytechnic Institute Troy, New York

Costandy S. Saba University of Dayton Research Institute Dayton, Ohio

Edward J. Myers Reynolds Metals Company Richmond, Virginia

William Scales Scales Air Compressor Corp. Carle Place, New York

Harold J. Noble Cincinnati Milacron Cincinnati, Ohio

Richard C. Schrama Dofasco, Inc. Operating Services Technology Hamilton, Ontario, Canada

Joseph M. Perez Pennsylvania State University University Park, Pennsylvania

Shirley E. Schwartz GM Research and Development Center Warren, Michigan

Douglas G. Placek FMC Corporation Princeton, New Jersey

Ronald L. Shubkin Albemarle Corporation Baton Rouge, Louisiana

Jack Poley Lubricon Beech Grove, Indiana

James H. Simpson, III Vogel Lubrication, Inc. Newport News, Virginia

Thomas E. Rajewski CPI Engineering Services Midland, Michigan

Donald J. Smolenski GM Research and Development Center Warren, Michigan

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John S. Straiton Marketing Technical Services Exxon Houston, Texas Robert L. Thom National Aeronautics and Space Administration Marshall Space Flight Center Alabama Douglas Vallance Cincinnati Milacron Cincinnati, Ohio.

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Frederick J. Villforth, III Texaco Research and Development Beacon, New York Donald F. Wilcock Tribolock Inc. Slingerlands, New York Zunde Yang Dept. of Materials Science and Engineering Northwestern University Evanston, Illinois

I Lubricant Properties

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1

Typical Lubricating Oil Properties Chun-I Chen

CONTENTS Typical Property Values of Lubricating Oils.........................................................................................3 Relations for Extending Physical Property Data................................................................................24 References...................................................................................................................................................32 Appendix.....................................................................................................................................................33

There are two components in this section. The first presents typical lubricating oil properties in table format. The second gives correlations that can be used to estimate properties under other conditions which are not included in the tables.

TYPICAL PROPERTY VALUES OF LUBRICATING OILS The data presented in this section are condensed from thousands of product records collected from manufacturers and data available in the literature. Typical values of most available properties are presented in Table 1. Additional tribological performance data are presented in Table 2. Mineral oils are grouped by application. Synthetic oils are categorized by functional group. The ID number for each record is provided for cross reference between Table 1 and Table 2. Records with the same ID number represent the same oil. The “Type of Oil” column groups oils in terms of application. Within an application group, lubrnicants of different specification are listed. The major property defined in a specification is viscosity. For major applications, such as gear oils, the specific viscosity grading system is used. For applications without a specific viscosity grading system, the ISO (International Standards Organization) viscosity grading system is used. The specification for engine oil is a combination of SAE (Society of Automotive Engineers) viscosity grade and American Petroleum Institute (API) service grade. The description in the second column of Table 1 is an abbreviation with two parts. The first part is a number or two numbers with a “W” in between that closely resembles the SAE grade. The second part consists of one or more letters that indicate API service grade. Following are the abbreviations and their meanings. Properties of motor oils for the SH service grade will generally match closely those with a “G” designation, as is also the case for CG commercial service oil properties closely matching those with an “F’ designation. G D D2 E F

API service grade SG API service grade CD API service grade CDII API service grade CE API service grade CF

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In most cases, the specification determines the formulation. If lubricants that meet the same specification have roughly the same properties, only one set of typical values is given. Sometimes, mainly due to the different base oil types, oils may meet the same specification but have very different values in some properties. In such cases, two or more sets of values may be listed. For example, both naphthenic mineral oils and paraffinic mineral oils can be formulated to meet the AGMA #1 specification; however, they have very different viscosity-temperature properties. Without a VI improver, paraffinic oils will show VI values around 100, whereas naphthenic oils have typical values of less than 50. The “AVG MW” column presents the average molecular weight of the lubricant oils. The “Common Additives” column lists the additives commonly found in oils that meet the given specification. The meanings of the abbreviations used in this column are listed as follows: C, anticorrosion; O, antioxidant; R, antirust; W, antiwear; F, defoam; DT, detergent; DS, dispersant; EP, extreme pressure; FM, friction modifier; MD, metal deactivator; P, pour depressant; VI, viscosity index improver.

RELATIONS FOR EXTENDING PHYSICAL PROPERTY DATA VISCOSITY-TEMPERATURE RELATIONSHIP ASTM method D341 can be used to obtain the viscosity-temperature relationship. A simplified form can be used to calculate the kinematic viscosity:

where: log v T A and B

= logarithm to base 10 = kinematic viscosity, [cSt] or [mm2/s] = Temperature, [K] or [°R] = dimensionless constants

The constants A and B can be evaluated for a fluid from two data points at different temperatures. Kinematic viscosities or temperatures for other points can then be readily calculated. To use this equation, the kinematic viscosity should be larger than 2.0 centistokes. For lower viscosities, the constant 0.7 increases according to relations in ASTM D341. ASTM has available “Standard ViscosityTemperature Charts” which give a straight line relation for any oil represented by Equation 1. The “Middle Range” chart of Figure 1 covers the temperature range of -40 to 150°C. Other charts extend from -70 to 370°C for viscosities from 0.18 to 20 million cSt. The viscosity index (VI) can be calculated by ASTM method D2270 and can be approximated from Figure 2. This arbitrary measure gives a relative viscosity-temperature sensitivity, with 100 representing little change in viscosity with temperature (the best paraffinic mineral oils in 1929) and 0 representing a great change (the poorest naphthenic oils in 1929). Note that the kinematic viscosity is related to dynamic viscosity as follows:

VISCOSITY-PRESSURE COEFFICIENT The viscosity-pressure coefficient, α, is defined as: Copyright © 1997 CRC Press, LLC.

FIGURE 1 “Middle Range” Standard Viscosity-Temperature Chart. From American Society for Testing and Materials, Philadelphia, PA. With permission. Copyright © 1997 CRC Press, LLC.

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FIGURE 2 Viscosity index from kinematic viscoties at 40°C and 100°C. The So correlation24,25 can be used to calculate the viscosity-pressure coefficient.

where: α m0 ν0 r

= pressure-viscosity coefficient, [10-8Pa-1] = viscosity-temperature property from the ASTM Walther equation and equal to (ASTM slope)/0.2; ASTM slope = constant B in Equation 1 = atmospheric kinematic viscosity at the temperature of interest, [mm2/s] = atmospheric density at the temperature of interest, [103kg/m3]

Note: kinematic viscosity at 1 atm and the temperature of interest should be above 1.0. Copyright © 1997 CRC Press, LLC.

Typical Lubricating Oil Properties

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Kinematic viscosity at pressure P and the temperature of interest can be obtained by the following equation. Significant variation in this viscosity-pressure coefficient can be expected when the pressure ranges above about 50–300 atm. (0.5–3 MPa).

where: ν P Po

= kinematic viscosity at pressure P and at the temperature of interest, [mm2/s] = pressure, [GPa] = atmospheric pressure (0.0001 GPa)

AVERAGE MOLECULAR WEIGHT Equation (2B2.3-1) of the API Technical Data Book26 is used to estimate the molecular weight of heavy petroleum fractions:

where: M ν100 ν210 sp gr

= molecular weight of petroleum fraction = kinematic viscosity of petroleum fraction at 100°F (37.8°C), [cSt] = kinematic viscosity of petroleum fraction at 210°F (98.9°C), [cSt] = specific gravity, 60/60°F (15.6/15.6°C)

Alternately, Equation 2B2.1-1 of the API Technical Data Book26 can also be used.

where: Tb sp gr

= mean average boiling point of petroleum fraction, [°R] = specific gravity, 60/60°F (15.6/15.6°C)

THERMAL EXPANSION COEFFICIENT The correlation from Table 6 of the “Petroleum Measurement Tables” of the API Manual of Petroleum Measurement Standards is used to estimate the thermal expansion coefficient:

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where: ρT ρ60F T

= density at the temperature T, [103kg/m3] = density at 60°F, [103kg/m3] = temperature, [°F]

THERMAL CONDUCTIVITY Equation 12A3.1-1 of the API Technical Data Book26 is used to estimate the thermal conductivity of petroleum fraction liquids:

where: k T

= thermal conductivity, [Btu/h/ft/F]. (1 Btu/h/ft/F = 1.7307 W/m/K.) = temperature, [°F]

Another equation27 can also be used:

where: k T sp gr

= thermal conductivity, [Btu in./(h ft2oF)]. (1 Btu in./(h ft2oF) = 0.1442279 W/m/K.) = temperature, [°F] = specific gravity, 60/60°F (15.6/15.6°C)

SPECIFIC HEAT Equation 7D2.2-1 of the API Technical Data Book26 is used to estimate the specific heat of petroleum fraction liquids:

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Typical Lubricating Oil Properties

sp gr Tb

29

= specific gravity 60/60°F (15.6/15.6°C) = mean average boiling point of petroleum fraction, [°R]

If the kinematic viscosities at 100°F and 210°F are known, then Equation 3 can be used to estimate the molecular weight. Equation 4 can then be used to estimate Tb. Another equation27 can also be used.

where: Cp sp gr T

= specific heat, [Btu/Ib/°R] (1 Btu/Ib/°R = 4186.8 J/kg/K.) = specific gravity 60/60°F (15.6/15.6°C) = temperature, [°F]

BULK MODULUS The Song correlation28 is used to estimate the bulk modulus. Only the values at 40°C and 1 atm are given in Table 1. The modulus estimated by this correlation is isothermal secant bulk modulus, which is defined as:

where: B V0,T P P0 VP,T

= isothermal secant bulk modulus at pressure P and temperature T, [GPa] = specific volume at pressure P0and temperature T, [cm3/g] = pressure, [GPa] = atmospheric pressure (0.0001 GPa) = specific volume at pressure P and temperature T, [cm3/g]

The following equations and table are used in the correlation: (a) Mineral oil lubricants and pure hydrocarbon

where: BP,T

= isothermal secant bulk modulus of mineral oil and pure hydrocarbon at pressure P and temperature T, [GPa] B0,T = isothermal secant bulk modulus at atmospheric pressure and temperture T, [GPa] log10(B0,T)= 0.3766*[log10(v0,T)]0.3307-0.2766 ν0,T = kinematic viscosity at atmospheric pressure and temperature T, [cSt] AT = -0.01382T + 5.851 T = temperature, [°CJ P = pressure, [GPa]

(b) Nonmineral oil based fluid Copyright © 1997 CRC Press, LLC.

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where: B′P,T BP,T S

= isothermal secant bulk modulus of nonmineral-based fluid at pressure P and temperature T, [GPa] = isothermal secant bulk modulus from Equation 9, [GPa] = deviation factor for nonmineral-based fluids from the following table, [GPa] Fluid Class Methyl silicone Phenyl silicone Perfluoropolyether Polybutene Poly(a-olefin) Ester Pentaerythritol ester Phosphate ester Polyphenyl ether

Deviation Factor S [GPa] -0.755 -0.160 -0.823 -0.268 -0.091 +0.092 +0.219 +0.301 +0.709

Once the bulk modulus at the temperature and pressure of interest is obtained by Equation 9 or 10, Equation 8 can then be used to calculate the specific volume or density at that temperature and pressure. FIRE AND FLASH POINT The COC flash point (ASTM D92) is given for all of the liquid lubricants in the database. The fire points (ASTM D92), on the other hand, are not generally available for many lubricant formulations. The fire point is equal to 1.1 times the flash point for single fractions from vacuum distillation. When several fractions of mineral oil are blended to produce a finished lubricant, the fire point is higher than 1.1 times the flash point. Since the same finished product, e.g., a 10W-30 motor oil, can be prepared from either a single mineral oil fraction or several mineral oil fractions, a supplier usually gives only the flash point. In Table 1, the fire point has been established using the relationship for a single mineral oil fraction. This means that the fire point of products prepared from blends of mineral oil fractions will be somewhat higher than those given in the table. The flash and fire points are given in Table 1 to the closest 5°C. CLOUD AND POUR POINT The cloud point is the temperature at which the oil becomes cloudy, indicating the formation of a second insoluble phase, usually due to wax formation as described in ASTM D2500. The pour point is the temperature at which the oil ceases to flow as described in ASTM test procedure D97. The lowtemperature use of a fluid is primarily limited by the pour point, which can be influenced by the use of pour depressants. For a given pour point, the cloud point may vary considerably due to the completeness of the dewaxing and the effectiveness of the pour depressant. Since the cloud point is not directly involved in low-temperature pumpability or flow tests and is not generally available for finished lubricant formulations, no cloud point values are presented in the database. GAS SOLUBILITY A semiempirical correlation proposed by Beerbower (Beerbower, 1980) can be used to estimate the solubility of gases at atmospheric pressure.

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Liquid solubility parameter, S1, is approximately 18.0 for diesters commonly used in aircraft fluids, 18.5 to 19.0 for higher esters, 18.41 for methyl phenyl silicone, 15.14 for dimethyl silicone, 18.29 for tri-2-ethylhexyl phosphate, and 18.82 for tricresyl phosphate. Some gas solubility parameters, S2, are given in the following table.29

Effect of dissolved gases on viscosity of lubricants at 1000 psi (7.0 MPa) is shown in the following table.29

VAPOR PRESSURE OF LUBRICATING OILS Some vapor pressure data are shown in the following table.23 Copyright © 1997 CRC Press, LLC.

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REFERENCES 1. Exxon Product Summary, Exxon Company, Houston, TX (1990). 2. 1991 Lubricating Oil, Grease and Antifreeze/Coolant Digest, Texaco Lubricants Company North America, Houston, TX (1991). 3. Product Salesfax Digest, Chevron Research and Technology Company, Richmond, CA (1991). 4. Mobil Brief Product Descriptions, Mobil Oil Corporation, Fairfax, VA (1989). 5. Product Guide, Conoco Inc., Houston, TX (1991). 6. Specific data sheets from Pennzoil Company, Houston, TX (1991). 7. GE Silicones, General Electric Company, Waterford, NY (1991). 8. Information about Silicone Fluids, Dow Corning Corporation, Midland, MI. 9. Information about MOLYKOTE® Specialty Lubricants, Dow Corning Corporation, Midland, MI. 10. Polyphenyl ether data from Monsanto, St. Louis, MO. 11. The Polyglycol Handbook, The Dow Chemical Company, Midland, MI (1988). 12. UCON Fluids & Lubricants, Union Carbide Corporation, Danbury, CT (1987). 13. Durad Lubricant Additives, FMC Corporation, Philadelphia, PA. 14. KRYTOX General Purpose Lubricants, Du Pont Company, Wilmington, DE (1988). 15. KRYTOX Fluorinated Oils, Du Pont Company, Wilmington, DE (1988). 16. List of Specifications for Du Pont Code Lubricants, Du Pont Company, Wilmington, DE (1986). 17. Product Information, Keystone Lubricants Division, Elf Lubricants North America, Inc., King of Prussia, PA. 18. Lubricants for Switches and Electric Contacts, William F. NYE, Inc., New Bedford, MA. 19. Booser, E. R., Ed., CRC Handbook of Lubrication, Vol. 1 (1983) and Vol. 2 (1984), CRC Press, Boca Raton, FL. 20. 1990 Annual Book of ASTM Standards, Section 5, Vols. 1–3, ASTM, Philadelphia, PA (1990). 21. MIL-L-7808, Standardization Documents Order Desk, Building 4D, 700 Robbins Avenue, Philadelphia, PA 19111-5094. 22. MIL-L-23699, Standardization Documents Order Desk, Building 4D, 700 Robbins Avenue, Philadelphia, PA 19111-5094. 23. Tables of Useful Information, Lubetext DG-400, Exxon Corp., Houston, TX (1992). 24. So, B. Y. C. and Klaus, E. E., Viscosity-pressure correlation of liquid, ASLE Trans., 23, 4, 409 (1980). 25. Wu, C. S., Klaus, E. E., and Duda, J. L., Development of a method for the prediction of pressureviscosity coefficients of lubricating oils based on free-volume theory, J. Tribol., 111, 121 (1989). 26. Daubert, T. E. and Danner, R. P., Technical Data Book — Petroleum Refining, 4th ed., American Petroleum Institute, Washington, D.C. (1992). 27. Cragoe, C. S., Thermal Properties of Petroleum Products, Miscellaneous Publication of the Bureau of Standards No. 97, Washington, D.C, (1929). 28. Song, H. S., Klaus, E. E., and Duda, J. L., Prediction of bulk moduli for mineral oil based lubricants, polymer solutions, and several classes of synthetic fluids, J. Tribol., 113, 675 (1991). 29. Klaus, E. E. and Tewksbury, E. J., Liquid lubricants, in CRC Handbook of Lubrication, vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 242. 30. Beerbower, A., Estimating the solubility of gases in petroleum and synthetic lubricants, ASLE Trans., 23, 335 (1980). Copyright © 1997 CRC Press, LLC.

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APPENDIX GENERAL DESCRIPTION OF THE GROUPING CATEGORIES Adhesive Oil: This category includes lubricants used in machine slideways and chain oils. Lubricants in this category contain tackiness additives to improve the adhesiveness. Compressor Oil, Food Grade: Lubricants in this category can be used in the environment where there is the possibility of incidental contact with food. Compressor Oil, Other: All other compressor oils except food grade oils. Engine Oil, Piston, Automotive: Engine oils used for automotive piston engines including gasoline and diesel engines. Engine Oil, Piston, Diesel: Engine oils used specifically for diesel engines. Engine Oil, Piston, Aircraft: Engine oils used in aircraft piston engines. Engine Oil, Piston, Gas: Engine oils used in piston engines fueled by natural gas or propane. Engine Oil, Piston, Marine: Diesel engine oils used in marine applications. Engine Oil, Piston, 2 Cycle: Engine oils for two-cycle piston engines. Engine Oil, Gas Turbine, Aircraft: Engine oils for aircraft gas turbine engines. Engine Oil, Gas Turbine, Land-Based: Engine oils for land-based gas turbine engines. Flushing Oil: This category includes petroleum base oils and once-through oils. They are prepared with no additives or with additives that produce a protective film after flushing. Gear Oil, AGMA R&O: Rust and oxidation inhibited gear oils. Gear Oil, AGMA EP: Gear oils that contain EP additives in addition to rust and oxidation inhibitors Gear Oil, AGMA Comp.: Gear oils with rust and oxidation inhibitors and friction modifiers. Heat Transfer Oil: Oils used in heat transfer operations. Hydraulic Oil: Oils used in hydraulic systems. They are usually formulated for good oxidation stability and lubricity. Machine Oil (Utility Oil): General purpose lubricants. Paper Machine Oil: Circulating oils in paper machines. Refrigeration Oil: Refrigeration compressor lubricants. Steam Turbine Oil: Circulating oils used in steam turbine systems. Textile Oil: Lubricants for textile machinery parts. Transformer Oil: Oils used in circuit breakers, switches, transformers, and other electrical apparatus for insulating, cooling, or both. Transmission Oil, ATF: Automatic transmission fluids. Transmission Oil, Automotive Gear Oil: Automotive manual transmission fluids. Transmission Oil, Tractor Hydraulic Fluid: Fluids used in transmissions, final drives, wet brakes, and hydraulic systems of tractors.

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Comparison of Properties of Synthetic Fludis* Wilfried J. Bartz

* Supplied by Prof. Wilfred J. Bartz.

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Polyalphaolefins Ronald L. Shubkin

Polyalphaolefin (PAO) fluids are synthetic hydrocarbons designed to provide superior lubrication performance over a wide temperature operating range.1 They are manufactured by a two-step process from linear alpha-olefins, which are themselves produced from ethylene. The first synthesis step entails oligomerization, which simply means a polymerization to relatively low molecular weight products. alpha-Olefin → Dimer + Trimer + Tetramer + Pentamer, etc. For the production of low-viscosity (2 to 10 cSt @ 100°C) PAO fluids, the catalyst for the oligomerization reaction is boron trifluoride. The BF3 catalyst is used in conjunction with a protic co-catalyst such as water, an alcohol, or a weak carboxylic acid. The BF3· ROH catalyst system is unique because of its ability to form highly branched products with the oligomer distribution peaking at the trimer. High-viscosity (40 and 100 cSt @ 100°C) PAO fluids are manufactured using Ziegler-Natta catalysts such as alkylaluminum compounds in conjunction with organic halides. The second step in the manufacturing process is hydrogenation of the unsaturated oligomer to enhance chemical inertness and oxidative stability. The reaction is carried out over a metal catalyst such as nickel or palladium. Distillation of the reaction mass to give specific product cuts may be done before or after hydrogenation. One distinct advantage in the manufacture of PAO fluids is that they can be “tailormade” to fit end-use requirements by manipulation of reaction variables which include:2 • Chain length of olefin raw material • Temperature • Time • Catalyst and co-catalyst type and concentration • Distillation of final product. Although choice of starting olefin can exert a major influence on product properties, commercial PAOs are generally derived from 1-decene to provide the broadest range of operational temperature. The following tables list properties of various grades of commercial decene-derived PAO base fluids.3 The widely used convention for designating the PAO grade is to use the kinematic viscosity (KV) in centistokes (cSt) at 100°C. In other words, a PAO fluid with KV100°C of approximately 2 cSt would be referred to as PAO 2. Differences in properties among the various grades illustrate what can be accomplished by manipulation of the reaction parameters. Some products are coproduced and separated by distillation. The properties listed in Tables 1 to 5 are typical of base fluids currently available; they do not represent the specifications of any particular producer and do not include property enhancements that can be obtained by the proper choice of additives. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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REFERENCES 1. Shubkin, R. L., Polyalphaolefins, Synthetic Lubricants and High-Performance Functional Fluids Shubkin, R. L., Ed., Marcel Dekker, New York, 1992, 1–40. 2. Shubkin, R. L. and Kerkemeyer, M. E., Tailor Making PAOs, 7th Int. Colloq. Automotive Lubrication, Technische Akademie Esslingen, Federal Republic of Germany, January 16–18, 1990; also, J. Synth. Lubr., 8(2): 115–134, 1991. 3. Shubkin, R. L., Polyalphaolefins, CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 219–252.

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Organic Esters Bruce J. Beimesch

The investigation of synthetic esters as lubricants goes back to the 1930s. From that original foundation, extensive research has led to the commercialization of numerous esters that find utility in almost every area of lubrication. Applications include the lubrication of sump systems such as automotive gasoline and diesel engines, car and truck axles and transmissions, industrial gear boxes, oven chain and conveyer systems, both air and water cooled 2-stroke engines, turbine aircraft and stationary engines, hydraulic systems, and many other applications. Synthetic esters have been utilized in synthetic greases for both high and low temperature applications since the 1950s. Military specifications for synthetic greases demand torque testing at -100°F. There are many types of organic esters suitable as lubricants and these include:

Diesters Polyol esters Dimer esters Aromatic esters Monoesters Esters are the reaction products of organic acid and an organic base. Organic Acid + Organic Base → Ester + Water The organic carboxylic acids are derived from natural and petrochemical processes. They can be monofunctional and polyfunctional. The organic alcohols, derived mainly from petrochemical processes, likewise can be simple alcohols or polyfunctional polyol alcohols such as pentaerythritol, neopentyl glycol, and trimethylolpropane. The properties of the ester can be controlled or tailored by the choice of organic acid and alcohol used in the esterification. The following tables list over 40 of the most common commercially available synthetic ester basestocks. The common properties are listed and should be useful to the lubricant formulator as well as design engineers. The properties are typical of commercial material. There may be some variation in the same ester, but these will be minor. Esters also respond favorably to additives used to enhance performance, such as antioxidants, corrosion inhibitors, anti-wear agents, and load-carrying additives. Some data are included on biodegradability that utilized the CEC L-33-A-94 test method. In general esters are readily biodegradable, and this feature is important to formulators who are developing ecologically compatible lubricants.

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FIGURE 1 Lubricants, synthetic oils: thermal conductivity vs. temperature, —, MIL-L-7808, —-, MIL-L- 23699. (From Data Book for Designers, Exxon Company, Houston, TX, 1973. With permission.)

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FIGURE 2 Thermal expansion, expansion factor vs. temperature, —, Ester lubricants; —-, petroleum lubricants. Multiply volume at 60°F by Expansion Factor to obtain approximate volume at desired temperature. (From Data Book for Designers, Exxon Company, Houston, TX, 1973. With permission.)

FIGURE 3 Lubricants, synthetic oils: specific heat vs. temperature. Solid line, diester; dashed line, petroleum. (From Data Book for Designers, Exxon Company, Houston, TX, 1973. With permission.)

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REFERENCES 1. Gunderson, R. C. and Hart, A. W., Synthetic lubricants, Reinhold Publishing, New York, 1962. 2. Shubkin, R. L., Synthetic lubricants and high-performance functional fluids, Marcel Dekker, New York, 1993. 3. Perez, J. M. and Klaus, E. E., Dibasic acid and polyol esters, in CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 237–252. 4. Data Book for Designers, Exxon Company, Houston, TX, 1973.

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Polyalkylene Glycols William L. Brown

CONTENTS Structure............................................................................................................................................49 Nomenclature...................................................................................................................................49 Sequencing.........................................................................................................................................52 References.........................................................................................................................................59

STRUCTURE Polyalkylene glycols (PAGs) are usually made from the reaction of ethylene and/or propylene oxides with a nucleophilic starter, such as an alcohol. They can be represented by the following structure:

R R´ R´´

= H or alkyl group = H, CH3, or alkyl group = H or alkyl group

NOMENCLATURE The polyalkylene glycols described in this chapter use the following nomenclature: # - AX - **** - Y # A X **** Y

Weight percentage of polymerized EO in the PAG; the remainder of the monomer feed is PO. O for oil soluble; W for water soluble at 20°C Represents the starter alcohol. M = methanol; B = butanol; D = low molecular weight diol such as ethylene glycol, diethylene glycol, propylene glycol, or dipropylene glycol. Average molecular weight of the polyalkylene glycol. Represents the monomer sequencing. R = random; Bn = normal block sequencing, (EO)X- (PO)y- (EO)X; Br = reverse block sequencing, (PO)X(EO)X- (PO)X. Polymers made from 100% propylene oxide are not followed by a monomer sequencing letter.

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SEQUENCING The sequencing of the oxide monomers can be random or blocked: random PAG = M-ABBAABAAABABABABB or BAABABBBAA-D-BBBABAAB blocked PAG = AAAAABBBB-D-BBBBAAAAAA or M-BBBBBBBBBAAAAAAAAA M is a monofunctional alcohol starter D is a diol starter A, B are alkylene oxide monomers Copyright © 1997 CRC Press, LLC.

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FIGURE 1 Viscosity of assorted oil-soluble polyalkylene glycols as a function of temperature. (From UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1979. With permission.)

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FIGURE 2 Viscosity of assorted water soluble polyalkylene glycols as a function of temperature. (From UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1979. With permission.)

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FIGURE 3 The viscosity of aqueous solutions of assorted polyalkylene glycols. (From UCON) Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1987. With permission.)

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FIGURE 4 The FDA status of various random polyalkylene glycols. (From UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1992. With permission.)

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REFERENCES 1. Technical literature, UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1987. 2. Technical literature, Technical Data on Pluronic Polyols, BASF Wyandotte Corp., MI, 1978. 3. Technical literature, Technical Data on Plurionic Nonionic Surface Active Agents, BASF Wyandotte Corp., 1978. 4. Technical literature, UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1992. 5. Technical literature, UCON Fluids & Lubricants, Union Carbide Corp., Tarrytown, NY, 1979.

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Phosphate Esters D. G. Placek and M. P. Marino

Phosphate esters were first introduced as antiwear additives for lubricants in the 1930s. Original applications included automotive crankcase lubricants, gear oils, and aircraft engine oils. Phosphate esters were identified as less flammable hydraulic fluid base stocks in the 1940s, increasing the safety of new aircraft designs that made increasing use of hydraulic control systems. This section will present the physical and performance properties of commercially available phosphate esters used as functional fluid base stocks, lubricant base stocks, and lubricant antiwear additives. Neutral, trisubstituted, or tertiary esters of orthophosphoric acid are represented by the general structure:

Phosphate esters are commercially available as trialkyl, alkyl/aryl, or triaryl phosphates. Trialkyl phosphates are symmetrical (R’ = R” = R’”), and are based on straight or branched alkyl groups which contain from 2 to 8 carbons. The first commercialy significant triaryl phosphate esters produced were tricresyl phosphate (TCP) and trixylenyl phosphate (TXP), which are referred to as “natural” phosphate esters because the cresols and xylenols used in their manufacture are derived from petroleum oil or coal tar extracts. “Synthetic” analogs of the natural phosphate esters were developed in the 1960s in order to overcome raw material availability concerns and lower product costs. Isopropylphenyl and tertiarybutylphenyl phosphate esters are now commercially available in a variety of viscosity grades. Typical chemical structures for several triaryl phosphates are shown below:

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The only alkyl/aryl phosphates which are currently used by the lubrication industry are dibutyl phenyl phosphate and isodecyl diphenyl phosphate. The physical property and performance data presented in this section are gathered from a variety of commercial literature and technical publication sources. The authors have attempted to present the most accurate and current data possible by selecting data points that are representative of the average value expected from a standard commercially available material.

ACKNOWLEDGMENT The authors wish to thank Mr. Jeff Kimak, Mr. Arvind Rao, and Ms. Victoria Baikova for their assistance in generating data and searching the literature to ensure the accuracy of this publication. We also wish to thank the Process Additives Division of FMC Corporation for supporting this effort.

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FIGURE 1 Viscosity/pressure relationship of IPPP/46 phosphate ester fluids.4

REFERENCES 1. M. P. Marino and D. G. Placek, Phosphate Esters, Booser, E. R., Ed., CRC Handbook of Lubrication and Tribology, Vol 3: Monitoring, Materials, Synthetic Lubricants and Applications, CRC Press, Boca Raton, FL, 1994, 269. 2. Marino, M. P., Phosphate esters, Shubkin, R. L., Ed., Synthetic Lubricants and High-Performance Functional Fluids, Marcel Dekker, New York, 1992. 3. Durad Lubricant Additives, FMC Corporation, Philadelphia, 1994. 4. Reolube HYD Fire Resistant Fluids, FMC Corporation, Trafford Park, Manchester, U.K., 1994. 5. Houghto-Safe 1000 Series Phosphate Ester Fluids, Houghton International Inc., Valley Forge, PA, 1990. 6. Fyrquel Fire Resistant Hydraulic Fluids, Tech. Bulletin 88–151, Akzo Chemicals Inc., Chicago, IL, 1988. 7. FMC Corporation, Philadelphia, unpublished data. 8. Schulz, W. W., Navratil, J. D., and Bess, T., Science and Technology of Tributyl Phosphate, CRC Press, Boca Raton, FL, 1989. 9. Hatton, R. E., Phosphate Esters, Gunderson, R. C. and Hart, A. W., Eds., Synthetic Lubricants, Reinhold, New York, 1962, chap. 4. 10. Sears, J. K. and Darby, J. R., The Technology of Plasticizers, SPE Monograph, J. Wiley & Sons, New York, 1982. 11. Phillips, W. D., Phosphate Esters, Totten, G., Ed., A Handbook of Hydraulic Fluid Technology, Marcel Dekker, New York, 1997. Copyright © 1997 CRC Press, LLC.

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Silicones E. D. Brown

CONTENTS Nature of Silicones..........................................................................................................................75 Dimethylpolysiloxane......................................................................................................................76 Modified Dimethylpolysiloxanes...................................................................................................76 Methylphenyl or Diphenyldimethylpolysiloxane.........................................................................78 Methylalkyl and Methylarylalkylpolysiloxanes.............................................................................78 Greases...............................................................................................................................................78 References.........................................................................................................................................79 Silicones have been on the market for over 50 years and have become industry standards in a wide variety of demanding applications. The fluids and greases continue to play a major role in many applications where nothing else will serve or where their superior performance justifies their high cost. In lubrication, for example, bearings lubricated with either silicone oils or greases commonly last 3 to 20 times longer at high temperatures than their organic counterparts.

NATURE OF SILICONES The term “silicone” is shorthand for the more descriptive designation “organopolysiloxane.” The silicone fluids are clear, water white, nontoxic materials, inert, tasteless and odorless. Since they are synthetic, built up from building block units, a nearly infinite number of combinations of molecular structure are possible. They consist of a chain skeleton of alternating silicon and oxygen atoms with various groups occupying the remainder of the valances (Figure 1). These are either organic groups or such functional moieties as hydrogen or hydroxy units.

FIGURE 1 Siloxane surface orientation. With relatively standard organochemical techniques, one can add a wide variety of side units to the basic silicon-oxygen chain, and substantial literature covers them. Practical considerations, however, have dictated that the following are those finding the most use. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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DIMETHYLPOLYSILOXANE This dimethyl family is by far the most common of the silicones. There are, for example, 20 standard viscosity grades, ranging from 5 to 600,000 cSt at 25°C. By use of standard blend charts (Figure 2) any intermediate viscosity fluid can be prepared. With the exception of pour point for fluids over 60,000 cSt (600,000 fluid has a pour point of -32°C), and volatility and flash point for fluids under 50 cSt, all properties other than viscosity are essentially the same for fluids of all viscosities. Among the many applications for these fluids, they are used in polishes, release agents, antifoam agents, damping fluids, and light duty lubricants — particularly for plastics.

FIGURE 2 Blend scale for intermediate viscosities. The most unique properties of these fluids include their viscosity-temperature characteristics, their low temperature fluidity, their resistance to shear breakdown, and their low surface tension. No other fluid has a smaller change in viscosity with temperature (Figure 3). Their pour points are below -55°C, and their surface properties as shown by their low Van der Waals forces are responsible for their outstanding release and antifoam characteristics.

MODIFIED DIMETHYLPOLYSILOXANES Addition of small amounts of phenyl, chlorophenyl, or branched-chain dimethyl dramatically lowers the pour point and makes these fluids useful under extreme conditions (Table 1). Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Viscosity/temperature relationships for dimethylsiloxanes. Centistokes viscosities at 25°C given at right. (From Demby, D. et al., in Synthetic Lubricants and High Performance Fluids, Shubkin, R. L., Ed., Marcel Dekker, New York, 1992, 183.)

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METHYLPHENYL OR DIPHENYLDIMETHYLPOLYSILOXANE Addition of relatively large amounts of phenyl improves both the oxidative stability and radiation resistance of these fluids. This stability leads to their use in high temperature greases and as snubber fluids in nuclear reactors. Their almost indefinitely long life at 260°C (500°F) comes, however, at the cost of losing some of the flat viscosity-temperature characteristics of the dimethyl polysiloxane (Figure 4).

FIGURE 4 Viscosity/temperature relationships for a number of siloxanes. (From Barnes, J. E. and Wright, J. H., 55th NLGI Annu. Meeting, Tampa, FL, 1988.)

METHYLALKYL AND METHYLARYLALKYLPOLYSILOXANES Longer chain alkyl groups and unsaturated aryl groups such as the styrenes can also be added to the basic silicon-oxygen skeleton. These groups produce a product which not only allows treated surfaces to be painted or soldered, but also protects metal combinations such as aluminum-steel where other lubricants fail.

GREASES Although all of the fluids can be made into greases, the most common are those made with phenyl-containing fluids. For the best lubrication, specialty greases are made with either the chlorophenyl or methyl alkyl fluids (Table 2).

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REFERENCES 1. Brown, E. D., Methyl alkyl silicones. A new class of lubricant, ASLE Trans., 9, 31, 1966. 2. Barnes, J. E. and Wright, J. H., Silicone greases and compounds: their components, properties, and applications, NLGI 55th Annual Meeting, Tampa, FL, 1988. 3. Hardman, B. and Torkelson, A., Silicones, John Wiley & Sons, New York, 1989. 4. Demby, D. H. et al., in Synthetic Lubricants and High Performance Fluids, Shubkin, R. L., Ed., Marcel Dekker, New York, 1992, 183. 5. Brown, E. D., in Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 305–321.

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Perf luoroalkylpolyethers Gregory A. Bell

INTRODUCTION Perfluoroalkylpolyether (PFPE) fluids are composed entirely of carbon, fluorine, and oxygen. They are colorless, odorless, completely inert to most chemical agents including oxygen, compatible with most other materials and liquid over a wide temperature range. Physical properties such as those given in Table 1 vary with molecular weight. Chemical properties and stability usually depend more on chemical structure than on molecular weight. The following four distinct types of PFPE oils are commercially available. Although all PFPE types exhibit similar physical and chemical properties, there are small and sometimes significant differences. PFPE-1 CF3CF2CF2O-[CF(CF3)CF2-O-]nCF2CF3 PFPE-2 CF3O-[CF(CF3)CF2-O-]y-[CF2-O-]mCF3 PFPE-3 CF3O-[CF2CF2-O-]z-[CF2-O-]pCF3 PFPE-4 CF3CF2CF2-O-[CF2CF2CF2-O-]qCF2CF3 Both PFPE-1 and PFPE-2 are nonlinear molecules because the polymer chains contain pendant trifluoromethyl groups, (-CF3). PFPE-4 and PFPE-3 contain no pendant groups and are linear. The linear PFPE structures show less change of viscosity with temperature and pressure when compared to nonlinear PFPE. PFPE-1 has a fully shielded polymer chain. PFPE-2 has a partially shielded polymer chain. PFPE-3 and PFPE-4 have nonshielded polymer chains and are thus not protected from acid-catalyzed cleavage.

REFERENCES 1. DelPesco, T. W., Perfluoroalkylpolyethers, in CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 287–303, 1994. 2. DelPesco, T. W., Perfluoroalkylpolyethers, in Synthetic Lubricants and High Performance Functional Fluids, Shubkin, R. L., Ed., Marcel Dekker, New York, 1992. Copyright © 1997 CRC Press, LLC.

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Polychlorotrifuoroethylenes Louis L. Ferstanding

Polychlorotrifluoroethylene (PCTFE) oils and greases are inert, nonflammable lubricants. Unlike hydrocarbon-containing lubricants, they are used with very aggressive oxidizing and acidic chemicals without danger of fire or explosion. The following property information is helpful in choosing which grade of inert oil to choose for specific applications. As the viscosity and molecular weight of the halocarbon oil grades increase, so do the density and pour and cloud points. Table 1 lists the viscosities and densities at several temperatures, as well as the pour and cloud points and the refractive indices. The grade numbers represent the viscosity in centistokes at 100°F. The high density of these oils (almost 2) means that the viscosity in centipoises is about twice the centistokes value; in hydrocarbon-based fluids (with densities of approximately 1), centiposie and centistokes values are about the same. When changing to a halocarbon oil the proper viscosity in centipoises should be used to determine the grade. Where the lubricating situation calls for an inert grease, the properties given in Table 2 will be helpful in choosing the appropriate grade. The viscosity of PCTFE oils change significantly with temperature. Figure 1 is a viscosity-temperature chart using ASTM D-341, showing the changes for the whole range of oils, which provides the information necessary to choose the appropriate oil grade. For some applications it is helpful to know the vapor pressure at operating temperature. Figure 2 gives typical pressures for all the grades. There are also PCTFE oils designed for pump use which have significantly lower vapor pressures. A variety of miscellaneous physical properties are compiled in Table 3.

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FIGURE 1 Typical viscosity vs. temperature of polychlorotrifluoroethylene oils. (Modified ASTM D-341.) Centipoise scale Copyright © 1997 LLC. assumes a fixed denCRC sity Press, of 1.92. (From Halocarbon® Inert L ubricants, Halocarbon Products Corp., 1996. With permission.)

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FIGURE 2 Typical vapor pressures for halocarbon oils. (From Halocarbon® Inert L ubricants, Halocarbon Products Corp., 1996. With permission.)

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10

Polybutenes John D. Fotheringham

Polybutenes are synthetic hydrocarbon polymers manufactured by the cationic polymerization of isobutene-rich raffinate streams available from sources that include naphtha steam cracking and refinery catalytic cracker operations. Polybutenes and liquid polymers available commercially in a wide range of viscosities from free flowing oils to viscous semisolid rubbers. The nomenclature used to define a particular grade of polybutene can be based on SSU viscosity (grade number = SSU viscosity at 100°C divided by 100) or reference to molecular weight. Grades representative of polybutenes available with typical property data are shown in Table 1. Viscosity and density against temperature relationships are provided in Figures 1 and 2. As products, polybutenes are essentially nontoxic, water white in color, and stable to light and air under normal storage conditions. High viscosity index and inherent tackiness are associated with polybutenes are relatively volatile compared to an equiviscous polyalphaolefin (PAO), and show high rates of evaporation. Volatile loss is much reduced as the molecular weight of the polybutene increases (Figure 3). At temperatures of between 250 and 275°C, polybutenes undergo a thermal depolymerization involving breakdown of the polymer chain into lower molecular weight hydrocarbons (Figure 4). This thermal route to decomposition occurs, leaving no carbon residue or staining on metal surfaces. Polybutenes are nonpolar and readily soluble in a range of organic solvents (Table 3). For lubricant applications, polybutenes are normally combined with other base fluids, and compatibility is found with all types of mineral oils, PAOs, alkyl benzenes, and most types of synthetic esters (Table 4). Examples of viscosity curves for combinations of polybutene with other base oils are shown in Figures 5 to 7. Polybutenes are used as a major component of 2-stroke engine oils (Table 5), high-pressure polyethylene compressor lubricants, metal-working lubricants, and speciality greases (Table 6) where the prime desire is for low smoke, low deposits or low toxicity and where volatility and oxidation resistance are less critical. At lower levels in formulations, the more viscous grades of polybutene are used for viscosity adjustment, viscosity index improvement, and to provide adhesiveness, tackiness, and resilience for greases and energy-efficient automotive and industrial oils. Polybutenes are a versatile class of polymer and impart a number of key performance benefits to these types of lubricants (Table 7). The typical chosen for use in the different oil and lubricants sectors are summarized in Table 8.

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FIGURE 1 Viscosity vs. temperature relationship for polybutenes.

FIGURE 2 Density vs. temperature relationship for polybutenes.

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FIGURE 3 Volatile loss of polybutene held at temperatures 100 to 180°C over a 10-h test period following ASTM D972 text method.

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FIGURE 4 Thermogravimetric analysis of polybutenes in air, showing volatile loss preceding polymer depolymerization at 250 to 275°C.

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FIGURE 5 Polybutene blended with solvent neutral 1 50 mineral oil at 100°C.

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FIGURE 6 Polybutene blended with 4 cSt polyalphaolefin at 40°C.

FIGURE 7 Polybutene blended with 13 cSt diester at 100°C.

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Vegetable Oils — Structure and 11 Performance Saurabh S. Lawate, Kasturi Lal, and Chor Huang CONTENTS Introduction....................................................................................................................................103 Composition...................................................................................................................................104 Structure/Performance Relationships........................................................................................104 Performance Properties................................................................................................................104 Applications....................................................................................................................................104 Acknowledgments..........................................................................................................................109 Notes and References....................................................................................................................115

INTRODUCTION Vegetable oils are found in the seed or fruit of various plants.1,2 They are predominantly triacylglycerols (triglycerides) in which three fatty acid groups are esterified to a glycerol backbone (Figure1).3

FIGURE 1 Structure of a triacylglycerol. Vegetable oils have to be extracted or expressed from the plant tissue in the “crude” form. The “crude” vegetable oil, although predominantly a triglyceride, contains several minor components like steroids, pigments, waxes, etc. In most applications these minor components and other impurities are removed, using specific purification steps. Purification involves one or all of the following steps: refining (free fatty acid removal), bleaching (color removal), deodorization (free acid and peroxide removal), winterization (wax removal).4 Copyright © 1997 CRC Press, LLC.

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COMPOSITION The physical and performance properties of a purified vegetable oil are largely influenced by the fatty acid groups present in the triacylglycerol (Figure 1). Common vegetable oils contain a combination of saturated and unsaturated fatty acids (Table 1). The double bonds in unsaturated fatty acids generally have the cis configuration. Some vegetable oils also contain significant amounts of fatty acid groups with chemical functionality, in addition to the saturated and unsaturated fatty acids (Table 2). Examples of vegetable oils containing special functionality are castor and lesquerella oils (hydroxy functionality),5 vernonia oil (epoxy function-ality),6 and meadowfoam oil (multiple double bonds separated by more than two carbon atoms).7Because of the presence of chemical functionality (Figure 5) these oils can be subjected to unique chemical modifications that can be exploited in lubricating applications.8-10 The chemical structures of various fatty acids found in vegetable oils are shown in Figures 2 to 5.

STRUCTURE/PERFORMANCE RELATIONSHIPS Qualitative generalizations can be made about the structure-performance relationships for common vegetable oils. Thus, it can be seen from Table 3 that oxidative stability increases with decreasing unsaturation, while the reverse holds true for low temperature properties. Because of this, the use of either highly saturated or highly polyunsaturated vegetable oils in lubricant applications is limited. In contrast to these extremes, a special class of oils, containing a high oleic content ( ≥75% oleic) and low polyunsaturated fatty acid content (linoleic and/or linolenic), displays good oxidative stability with acceptable low temperature properties.11,12 This makes them well suited for use in lubricants compared to conventional vegetable oils.

PERFORMANCE PROPERTIES Basic physical properties of various vegetable oils are shown in Table 4. General performance properties of interest in lubricating applications are shown in Table 5 while Table 6 shows friction and load bearing properties. Note that the data in Tables 4 to 6 are without additives. Table 7 shows the oxidative stability of vegetable oils in the presence of antioxidants.13 These data indicate how oxidative stability increases with decreasing unsaturation — decreasing iodine value.

APPLICATIONS Vegetable oils are mainly consumed in foods. However, they also serve as the primary feedstock for the oleochemical industry and are gaining popularity as lubricating base oils.14,15 Vegetable oils are obtained from renewable resources and are biodegradable. Thus they offer specific environmental benefits over mineral oil-based lubricants. This is significant in applications where the lubricant is “lost” in the environment, e.g., chain bar lubricants and hydraulic fluids for farm machinery.16 In addition to environmental benefits, vegetable oils also have certain performance advantages over conventional mineral oil base stocks. These include low volatility, high flash points, viscosity index, and excellent lubricity. The primary drawback of conventional vegetable oils is their lower oxidative stability relative to mineral oils and certain synthetic esters. However, with recent advances in breeding technology, it is becoming possible to alter the physical properties of conventional vegetable oils by changing fatty acid profiles. A specific example pertaining to lubricant applications is the improvement of oxidative stability by increasing the oleic content in various oils.17,18 Copyright © 1997 CRC Press, LLC.

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Copyright2 ©Structure 1997 CRCof Press, LLC. fatty acids. FIGURE saturated

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FIGURE 2 (continued).

FIGURE 3 Structures of monounsaturated fatty acids.

ACKNOWLEDGMENTS The authors would like to gratefully acknowledge the help provided by Mr. Rick Unger in per forming several fatty analyses shown in Table 1 and the Lubrizol Corporation for supporting this work. Copyright © 1997 CRC Press, LLC.

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FIGURE 4 Structures of polyunsaturated fatty acids.

FIGURE 5 Structures of fatty acids with special chemical functionality.

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FIGURE 5 (continued).

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NOTES AND REFERENCES 1. Olive oil is an example of a vegetable oil derived from a fruit. 2. Animal and dairy fats, although triglycerides, are not described in this monograph. Animal fats include tallow, lard, and chicken fat, and dairy fats include butter fat. The key feature in animal-derived oils is the presence of significant amounts of palmitoleic acid (3 to 4%), whereas in dairy fats it is the presence of small and medium chain fatty acid groups. 3. Jojoba oil is also not described in this monograph, since although it is commonly a vegetable oil, by source, it does not have the triglyceride structure; it is predominantly a simple monoester-oleyl oleate. “Jojoba Oil and Derivatives,” Wisniak, ].,Prog. Chem. Fats Other Lipids, 1977, 15, 167-218. 4. Bailey’s Industrial Oil and Fat Products, 4th ed., Vol. 2, D. Swern, Ed., Wiley-Interscience Publications, New York, 1982. 5. “The Triglyceride Composition, Structure, and Presence of Estolides in the Oils of Lesquerella and Related Species,” Hayes, D. G., Kleiman, R., and Phillips, B. S., J. Am. Oil Chem. Soc. 1995, 72,559-569. 6. “Vernonia galamensis, Potential New Crop Source of Epoxy Acid,” Perdue, R. E., Carlson, K. D., and Gilbert, M. G., Economic Botany 1986, 40, 54-68. 7. “Meadowfoam: New Source of Long-Chain Fatty Acids,” Purdy, R. H. and Craig, C. D., J. Am. Oil Chem. Soc, 1987, 64, 1493-1498. Copyright © 1997 CRC Press, LLC.

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8. “Triglyceride Oils Thickened with Estolides of Hydroxy-Containing Triglycerides,” Lawate, S. S., U.S. Patent 5,427,704, 1995. 9. “Oils Thickened with Estolides of Hydroxy Containing Triglycerides,” Lawate, S. S., U.S. Patent 5,458,795, 1995. 10. “Estolides of Hydroxy-Containing Triglycerides That Contain a Performance Additive,” Lawate, S. S., U.S. Patent 5,451,332, 1995. 11 “Sunflower Product and Methods of Their Production,” Fick, G. N., U.S. Patent 4,627,192, 1986. 12. “Novel Sunflower Products and Methods for Their Production,” Fick, G. N., U.S. Patent 4,743,402, 1988. 13 Results are shown with a combination of a phenolic and amine type antioxidant. (LZ7652 available from the Lubrizol Corporation, Cleveland, OH.) 14. “Industrial Uses of Agricultural Materials — Situation and Outlook Report,” U.S. Department of Agriculture, Economic Research Service, 1993, IUS-1. 15. “Natural Fats and Oils — Renewable Raw Materials For the Chemical Industry,”Angew. Chem., 1988, 27, 41-62. 16. “Varieties of Rapeseed Oil and Derived Products for Use in Fuels and Lubricants,” Harold, S. C, Lai, K., and Lawate, S. S., Presented at the 9th Int. Rapeseed Congress, Cambridge, U.K., 1995. 17. “Modification of Plant Lipid Synthesis,” Töpfer, R., Martini, N., and Schell, J., Science, 1995, 268, 681-686. 18. “The Genetic Improvement of Seed Oil,” Robbelen, G., Chem. and Ind., 1991, 713-716.

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12 Additives — Chemistry and Testing Syed Q. A. Rizvi Almost all lubricants contain chemical additives. This is because unformulated lubricants, or base oils, do not possess properties necessary to perform effectively in today’s equipment. Additives improve the lubricating ability of base oils either by enhancing the desirable properties already present or by adding new properties. Base oil may be petroleum, synthetic, or biological in origin. Petroleum-derived base oils currently account for about 97% of total lubricant production. The processes used in their manufacture include distillation, deasphalting, solvent extraction, solvent dewaxing, and finishing. These processes help isolate materials that have suitable boiling points and physical and chemical properties for use in formulating lubricants. Synthetic base stocks are manufactured through transformations of petroleum-derived organic chemicals. These base stocks, often more expensive, are the only choice for extremely demanding applications where mineral oils cannot be used because of their inherent limitations. Blended base stocks, mixtures of synthetic base stocks and mineral oils, are often used to benefit from superior low temperature properties, high flash points, and high viscosity indices (less change in viscosity with temperature) of synthetic base stocks at a lower cost. Base stocks of biological origin include vegetable oils and animal fats that are obtained from seeds, fruits, and animal tissue. Because of their highly biodegradable nature and nonpetroleum origin, their use is increasingly becoming important. For some industrial applications, such as metalworking, even water can be used as a carrier of additives. Quality and quantity of additives in the lubricant depend upon the nature of the base fluid and the lubricant’s intended use. Typical additive concentrations are given in Table 1. The performance package can make up to 30% of the lubricant’s total composition. Engine oils and automotive gear oils, which place a higher demand on the lubricant, generally require larger concentrations of additives than less-demanding applications such as industrial lubricants and metalworking fluids. Lubricant additives perform a number of diverse functions and can be broadly classified into chemically-inert and chemically active types. Chemically inert additives improve a lubricant’s physical properties and include emulsifiers, demulsifiers, pour point depressants, foam inhibitors, and viscosity modifiers. Chemically-active additives, on the other hand, chemically interact with metals to form protective films and with polar oxidation and degradation products to make them innocuous. Such additives include dispersants, detergents, antiwear and extreme pressure agents, oxidation inhibitors, and rust and corrosion inhibitors. Additive types that are used in various lubricant applications are shown in Table 2; their chemistry type and the manner in which they function are provided in Tables 3 and 4. The structures of different classes of additives along with methods of their synthesis are described in Tables 5 to 10. Both physical and performance characteristics impact a lubricant’s effectiveness. For automotive lubricants, physical characteristics are established by the Society of Automotive Engineers (SAE) and are specified in its classification systems. For other lubricants, these are defined by original equipment manufacturers and end-users. Performance requirements, or performance specifications, Copyright © 1997 CRC Press, LLC.

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define a lubricant’s ability to reduce friction, resist oxidation, minimize deposit formation, and prevent corrosion and wear. Organizations that help establish such specifications in the U.S. include the SAE, the American Petroleum Institute (API), the American Society of Testing and Materials (ASTM), U.S. original equipment manufacturers (OEMs), and the U.S. military. Table 11 lists typical performance requirements by application. Lubricant effectiveness is assessed by bench tests and full scale testing in the laboratory and sometimes in the field. The laboratory tests are accelerated tests in real world equipment that simulate actual service conditions. These use actual engines, transmissions, axles, hydraulic pumps, and so on, and are carried out under standard conditions and according to prescribed procedures. The goal is to ascertain that the lubricants meet the performance requirements established by various organizations. Tables 12 and 13 show the parameters measured by different tests. Once a lubricant meets all the requirements, it is ready to be marketed for factory-fill and retail use.

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Greases Richard C. Schrama

CONTENTS Introduction.....................................................................................................................................................138 Base Oil Selection...........................................................................................................................................138 Additives............................................................................................................................................................145 Grease Selection by Load and Speed.........................................................................................................145 Re-Lubrication Intervals................................................................................................................................147 Mixing of Greases..........................................................................................................................................150 Grease Troubleshooting................................................................................................................................151 Reference...........................................................................................................................................................155

INTRODUCTION Grease is a lubricating oil thickened with about 7 to 20% of a gelling agent or “soap” to give a consistency ranging from a semifluid to a solid. Of typical thickeners listed in Table 1, lithium soaps predominate with about 65% of the market. Aluminum complex soaps and clay thickeners follow with about 6% each. Table 2 gives a formulation guide for greases to meet various application requirements. Grease consistency, or hardness, is designated by the National Lubricating Grease Institute (NLGI) grades as given in Table 3. These are based on cone penetration distance under standardized conditions in the ASTM Test D217. Standard test methods to determine the properties of greases are summarized in Table 4. Sufficient thickening agent is commonly used to give the widely used grade 2 consistency. Softer grades using less thickener give easier pumping in centralized lubrication systems and improved feeding for gear drives, multiple-row roller bearings, and slow speed sliding. Grade 3 greases are used for better channelling and better mechanical stability in some prepacked ball bearings and in large machinery. Hard grade 6 brick greases are inserted as blocks in sleeve bearing boxes of paper mills and railway rolling stock. The typical properties for greases with the different soap types are given in Table 5. This table can be used as an application guideline to determine which soap type has the best properties for the application requirements from the lubricant.

BASE OIL SELECTION Minimum viscosity for oil, or for the base oil in case of greases, necessary to provide a full lubricant film for the rolling elements in a bearing is given in Figure 1. For a mean bearing diameter dm = 50 mm at 1000 rpm, for instance, the base oil should provide a minimum viscosity of 20 mm2/s (cSt) at the bearing operating temperature to give an oil film thickness sufficient to just cover the Copyright © 1997 CRC Press, LLC.

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surface roughness of the bearing elements. If the base oil viscosity at the bearing temperature is only 10 cSt, the viscosity ratio v/v, drops to 0.5 and Table 6 indicates boundary lubrication with wear and reduced fatigue life. Use of a grease with extreme pressure properties and a higher base oil viscosity should then be considered. To provide suitable lubrication over a wide range of applications, mineral base oil viscosity of 100 cSt to 130 cSt at 40°C is common. Viscous oils in the 200 to 600 cSt range are employed for boundary lubrication and extreme pressure conditions, with drawbacks in increased friction, high temperature rise, and noise. Lower viscosity oils down to 25 cSt and lower at 40°C would provide for operation at lower ambient temperatures, easier dispensing, and possibly higher bearing speeds. Figure 2 shows the effect of temperature on viscosity for the various ISO oil classifications. A comparison of the properties which are critical for the mineral oil types which are used as base oils in the manufacture of the grease is given in Table 7. These oils can be used with most of the soap types. Copyright © 1997 CRC Press, LLC.

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The broader temperature limits for greases formulated with synthetic oils are indicated in Table 8. Limited by their higher cost to specialty applications, the synthetic greases currently make up only about 1 to 2% of the total grease market.

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FIGURE 1 Minimum allowable viscosity for rolling bearing lubrication: vh kinematic viscosity; n, operating speed; dm, mean bearing diameter. (From Rolling Bearing Greases, No. 522.111.88, Kluber Lubrication. With permission.)

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FIGURE 2 Viscosity-temperature chart. Viscosity numbers per ISO 3348-1975 for oils having viscosity index of 95; approximate equivalent SAE viscosity grades shown in parentheses. (From Boehringer, R. H., Grease, in ASM Handbook, Vol. 18, ASM International, Metals Park, OH, 1992, 123-131. With permission.)

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ADDITIVES Table 9 shows the typical additives which are added to grease formulations for oxidation resistance. The various grease properties and the corresponding typical chemical additive are indicated in Table 10. By no means are these the only chemical formulations that can be used for each of the grease properties. Also, the concentration ranges stated are typical in the industry. Some grease formulations may require the chemicals added to the base oils and soaps to be outside of the ranges stated.

GREASE SELECTION BY LOAD AND SPEED Speed and load effects on the grease can be evaluated from Figure 3. The ratio P/C is a measure of the specific load, where C is the dynamic load rating and P is the equivalent dynamic bearing load, P = X × Fr + Y × Fa. Fr is the radial load and Fa is the axial load. The values for C, X, and Y are available in bearing catalogues. The speed index, n × dm is a measure for the grease stressing by the speed, where n is the rotational speed of the bearing and dm is the mean bearing diameter (arithmetic mean of the bore diameter d and the outside diameter D). The speed index is multipled by a factor ka, depending on the bearing type. This factor takes into account the various degrees of sliding friction in the bearings. Figure 3 is divided into three load ranges whose limits depend on the type of load and speed rating. Nearly all rolling bearing greases are suitable for the lubrication of rolling bearings operating under the load conditions of range I. Excluded are the greases with an extremely low or high Copyright © 1997 CRC Press, LLC.

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viscosity base oil, extremely stiff or soft greases, and some special greases, e.g., silicone greases which can only be used up to loads of P/C = 0.03. In the high speed and high load range, in the upper right hand corner of range I, higher operating temperatures necessitate the use of thermally stable greases resistant to temperatures significantly higher than the expected bearing operating temperature. In the case of high loads and low speeds, the use of solid lubricant additives will provide “chemical lubrication” or dry lubrication which replace the hydrodynamic lubrication where the lubricating film has been interrupted. Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Grease selection from the load ratio P/C and the relevant bearing speed index da n dm - Range I, Normal operating conditions. Range II, Range of heavy loads. Range II, High-speed range, calling for greases for high-speed bearings, ka values: ka = 1 deep groove ball bearings, angular contact ball bearings, four-point bearings, selfaligning ball bearings, radially loaded cylindrical roller bearings, thrust ball bearings. ka = 2 spherical roller bearings, tapered roller bearings, needle roller bearings ka = 3 axially loaded cylindrical roller bearings, full complement cylindrical roller bearings. (From the Lubrication of Roller Bearings, Pub. no. WL81115/3EA, FAG Bearings Corp., Danbury, CT. With permission.)

RE-LUBRICATION INTERVALS Grease replenishment or exchange is required if the service life of the grease is much shorter than the anticipated bearing life. Figure 4 shows the lubrication intervals as a function of the speed index of the bearings (kf × n × dm) for the failure probability of 10 to 20%. A kf value range is indicated for some bearing types. The higher values apply to the heavier series (higher load carrying capacity) and the smaller values to the lighter series of a bearing type. The lubrication intervals on Figure 4 apply to lubrication with a lithium soap base grease and temperatures up to 70°C, measured at the bearing outer ring, normal environmental conditions, and a mean bearing load corresponding to P/C< 0.1. Every rise in temperature by 15K over 70°C halves the lubrication interval of lithium soap-base greases with mineral base oil. Also, vibrations acting on the bearing reduce the lubrication intervals because they result in a separation of the grease into thickener and base oil. Contaminants, including water, penetrating through the seals also affect the lubrication intervals. With gap-type seals, an air current passing through the bearing considerably reduces the lubrication interval. The air current deteriorates the lubricant, carries oil or grease from the bearing, and conveys contaminants into it. For poor operating and environmental conditions, a reduced lubrication interval tfq is obtained from the equation:

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FIGURE 4 Lubrication intervals (grease service life) for bearings lubricated with lithium soapbase grease under favorable environmental conditions. Failure probability 10 to 20%. Reduced lubrication intervals tfqmust be taken into account for adverse operating and environmental conditions. The replenishment interval must be shorter than the lubrication interval (usual values: 0.5...0.7-tfq).

Table 11 shows the corresponding reduction factors for the above equation. An alternative method for determining the lubrication interval is given in Figure 5. This graphical method has been used by SKF and other bearing manufacturers for years. It is a much simpler way of obtaining a lubrication interval for each lubrication point in a piece of machinery. The amount of grease put into the bearing is critical. Usually, ball bearings are filled with grease to 1/3 of their free space. The grease fill quantity is determined by the following rule of thumb:

where

d B

= bearing bore = bearing width in mm

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FIGURE 5 Grease relubrication intervals: a, radical ball bearings; b, cylindrical or needle roller bearings; c, spherical or taper roller bearings and thrust ball bearings; d, bearing bore diameter. (From Boehringer, R. H., Grease, in ASM Handbook, Vol. 18, ASM International, Metals Park, OH, 1992, 123-131, and courtesy of SKF-USA, Inc. King of Prussia, PA. With permission.) Copyright © 1997 CRC Press, LLC.

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FIGURE 6 Calculation data for free space in roller bearings: a, curve; b, bearing type. (From Roller Bearing Greases, Pub. no. 522.111.88, Kluber Lubricated. With permission.) Figure 6 shows the free space for various width classifications in each bearing type. The correct grease fill quantity for rolling bearings is also determined by the bearing construction, speed, and sealing. A complete fill is permissible up to medium speeds, providing that the grease can escape into the bearing cavity or through a labyrinth seal. High speed bearings are filled up to 1/3 of the bearing free space. Excessive heat generation within the bearing is produced by over-greasing or by the incorrect choice of grease which is dynamically too viscous. The correct grease fill produces an initial temperature rise which is normal (Figure 7). After a short-term rise, the temperature drops back. The grease has then distributed itself throughout the bearing. Localized grease excesses would then have been displaced from the bearing tracks.

MIXING OF GREASES Rolling bearing greases with similar thickeners and chemically similar base oils can, on the whole, be mixed without a problem. It should be noted that some base oils may not be mixed with others and some thickeners are not compatible with one another. The rolling bearing grease then becomes either stiffer and hardens, or it becomes softer and flows from the bearing. Table 12 gives some indication as to the compatibility of petroleum greases using different types of thickeners. These comparisons are subject to the percentage of soap used in the grease and other additives in the grease formulation. The mixing of grease should, therefore, be avoided. If it is necessary to change to another grease to improve the lubricating properties, a complete cleanout of the bearing is to be recommended together with a fresh lubrication. If two greases can be mixed, the grease change can be Copyright © 1997 CRC Press, LLC.

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FIGURE 6 (continued), achieved by pumping the new grease through the bearing until the old grease and the grease mixture are extruded. Frequent relubrication with the new grease should then take place until the old grease or the grease mixture is removed from the bearing.

GREASE TROUBLESHOOTING The NLGI has set up a very comprehensive troubleshooting table for greases. Table 13 shows various applications (plain and rolling bearings, gears, sliding surfaces, universal joints, electric motors, couplings) and typical trouble symptoms, possible causes, and what to check for in the equipment with the problem. Copyright © 1997 CRC Press, LLC.

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FIGURE 7 Possible temperature sequence in rolling bearings. 1 = Over-lubricated and incorrectly grease lubricated; 2 = over-lubricated or incorrectly grease lubricated; 3 = normal temperature sequence. (From Roller Bearing Greases, Pub. no. 522.111.88, Kluber Lubrication. With permission.)

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REFERENCES 1. Boehringer, R. H., Grease, in ASM Handbook, Vol. 18, ASM International, Materials Park, OH, 1992, 123-131. 2. Lubricating Grease Guide, 3rd ed., National Lubricating Grease Institute, Kansas City, MO, 1994. 3.Klamann, D., Lubricants and Related Products, Verlag Chemie, Weinheim, Germany, 1984. 4. The Lubrication of Rolling Bearings, Pub. no. WL81115/2EA, FAG Bearings Corp., Danbury, CT, 1989. 5. Rolling Bearing Greases, Pub. no. 522.111.88, Kluber Lubrication North America, L.P., 1988. 6. CRC Handbook on Tribology, Vol. 1, 2, and 3, E. R. Booser, Ed., CRC Press, Boca Raton, FL, 1983, 1984, 1994. 7. Standard Handbook of Lubrication Engineering, McGraw-Hill Book Co., New York, 1968. 8. U.S. Steel Lubrication Engineers Manual, A.I.S.E. (Association of Iron and Steel Engineers), Pittsburgh, PA, 1996. 9. Boner, C. J., Manufacture and Application of Lubricating Greases, Reinhold Publishing, New York, 1954. 10. Annual Book of ASTM Standards, Volumes 5.01, 5.02, 5.03, Petroleum Products and Lubricants, revised annually, ASTM, Philadelphia.

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14 Solid Lubricants E. R. Booser CONTENTS Inorganic Compounds........................................................................................................................156 Organic Polymers.................................................................................................................................157 Metal Films............................................................................................................................................158 References.............................................................................................................................................158

These materials provide thin films between two surfaces to reduce friction and wear, generally for high temperatures, vacuum, nuclear radiation, aerospace, and other environments that preclude the use of conventional oils and greases. The wide range of solid lubricants can generally be classified as inorganic compounds, organic polymers, and metal films.1-3 The inorganic compounds and polymers are commonly used in a bonded coating over chemical conversion coatings (See Section VII, Chapter 57) to provide lower friction and wear on metal surfaces.4’5

INORGANIC COMPOUNDS Most important of this group listed in Table 1 are layer-lattice solids in which bonding between atoms in an individual layer is by strong covalent or ionic forces and those between layers are by relatively weak ionic forces. Molybdenum disulfide and graphite are the preferred choices; others which find occasional use are tungsten disulfide, tungsten diselenide, niobium diselenide, calcium chloride, calcium iodide, and graphite fluoride.6 Graphite is commonly used as a dry powder or as a dispersion in water, oils, greases, or solvents. The dispersions are used for lubricating tools, dies, and molds for metal working and metal forming; with oxygen equipment; and for conveyors and other high temperature industrial applications. In vacuum or in atmospheres where no moisture can be absorbed, mixing with cadmium oxide, MoS2, or organic binders will restore the film-forming ability of graphite.1 Oxidation by air commonly sets a limit of about 55O°C; above 100°C high friction may be encountered from water desorption. Molybdenum disulfide has replaced graphite in many applications for its independence from the need for adsorbed vapors in providing lubrication, superior load capacity, and more consistent properties.1 MIL-M-7866 covers the most common grade.5 Above 400°C, MoS2 is oxidized to the trioxide which may be abrasive. For temperatures above 550°C, classes of materials that have shown promise are oxides formed on nickel-base and cobalt-base superalloys, and fluorides such as CaF2, BaF2, LiF, and MgF2 applied as ceramic-bonded coatings, by fusion bonding, or as components of plasmasprayed composite coatings.6 Various soft inorganic without a layer-lattice structure also find some use as solid lubricants: lime as a carrier in wire drawing, talc and bentonite as fillers for grease in cable pulling, Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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zinc oxide in high load capacity greases, and milk of magnesia for bolts. Toxicity has led to diminished use of basic white lead and lead carbonate.

ORGANIC POLYMERS Various polymers provide self-lubricating properties when applied as thin films, as bearing materials, and as binders for lamellar solids.1,3,4,7 Coatings are typically applied in powder or dispersion form (in thicknesses from 25 µtm and upward) and then fused to the surface to provide lubricity, wear resistance, or release properties. PTFE is outstanding in this group and is effective from about -200 to 250°C in providing a coefficient of friction in the 0.03 to 0.1 range. Chapter 52, in Section VI, provides details on the properties of polymers and their composites. Copyright © 1997 CRC Press, LLC.

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METAL FILMS Table 2 lists soft metals which in many ways are ideal solid lubricants.8 They have low shear strength, can be bonded to metal substrates, have good lubricity, and offer high thermal conductivity. Metal films are applied by electroplating, evaporation, sputtering, or ion plating. Finding use as bolt lubricants are electroplated silver and copper, as are commercial formulations incorporating powders of nickel, silver, copper, and lead. Tin, zinc, copper, and silver coatings are used as lubricants in metalworking where use of lead has been eliminated for its toxicity.9 Silver films are used in a variety of sliding and rolling contacts in vacuum and at high temperatures since silver is unique in forming no alloys with steel and is soft at high temperatures. While gallium is above its melting point under most conditions, it has been effective when applied as a coating in vacuum.

REFERENCES 1. Lancaster, J. K., Solid lubricants, in CRC Handbook of Lubrication, Vol. 2, CRC Press, Boca Raton, FL, 1984, 269-299. 2. Booser, E. R., Lubrication and lubricants, Encyclopedia of Chemical Technology, 4th ed., Vol. 15, John Wiley & Sons, New York, 1995, 463-517. 3. New Directions for Solid Lubricants, in New Directions in Lubrication, Materials, Wear, and Surface Interactions, Loomis, W. R., Ed., Noyes Publications, Park Ridge, NJ, 1985, 631-733. 4. Gresham, R. M, Bonded solid film lubricants, in CRC Handbook of Tribology and Lubrication, Vol. 3, CRC Press, Boca Raton, FL, 1994, 167-181. 5. Lipp, L. C, Lubr. Eng., 32, 574-584, 1976. 6. Sliney, H. E., Solid lubricants, in ASM Handbook, Vol. 18, 1992, 113-122. 7. Jamison, W. E., Plastics and plastic matrix composites, in CRC Handbook of Tribology and Lubrication, Vol. 3, CRC Press, Boca Raton, FL, 1994, 121-147. 8. Peterson, M. B., Murray, S. F., and Florek, J. J., ASLE Trans., 2, 225-234, 1960. 9. Schey, J. A., Tribology in Metalworking: Friction, Lubrication and Wear, ASM, Metals Park, OH, 1983.

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of Liquids, Aqueous 15 Viscosity Solutions, and Liquid Metals* The absolute viscosity of some common liquids at temperatures between -25 and 100°C is given in this table. Values were derived by fitting experimental data to suitable expressions for the temperature dependence. The substances are arranged by molecular formula in the modified Hill order. All values are given in units of millipascal seconds (mPa s); this unit is identical to centipoise (cp). Viscosity values correspond to a nominal pressure of 1 atmosphere. If a value is given at a temperature above the normal boiling point, the applicable pressure is understood to be the vapor pressure of the liquid at that temperature. A few values are given at a temperature slightly below the normal freezing point; these refer to the supercooled liquid. The accuracy ranges from 1% in the best cases to 5 to 10% in the worst cases. Additional significant figures are included in the table to facilitate interpolation.

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VISCOSITY OF AQUEOUS SOLUTIONS This table gives the absolute viscosity of aqueous solutions of several common compounds as a function of concentration expressed in mass %. Viscosity values are in units of millipascal seconds (mPa s), which is equivalent to centipoises (cp).

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16 Gas Properties Donald F. Wilcock CONTENTS Viscosity...........................................................................................................................................169 Specific Volume..............................................................................................................................174 Mean Free Path..............................................................................................................................174 Speed of Sound..............................................................................................................................176 Nomenclature.................................................................................................................................176 References.......................................................................................................................................176

In designing gas bearings, viscosity is usually the property of prime interest. A number of other physical properties may also be required, however, and are described in this section. Chemical properties of any particular gas may influence mixing of the gas with fluids in the system, reactions with other gases, or reaction with bearings or other surfaces. The designer should, therefore, ascertain from other sources the chemical reactivity of the gas. Data in Table 1 are abstracted from an extensive listing of thermophysical properties of liquids and gases.1 The first three columns give the common name of the gas, its chemical formula, and its molecular weight. Column four gives the boiling point in K at a pressure of 760 mmHg or 1.01 bar. Also given are specific volume in m3/kg, heat capacity Cp in kJ/kgK, speed of sound in m/s, viscosity in Pa-s, and the viscosity-temperature exponent in Equation 1.

VISCOSITY In addition to viscosity data listed in Table 1. Table 2 lists viscosities over a range from 100 K to 600 K for many gases.2 The viscosity of a gas is nearly independent of pressure over a wide range of lower pressures, but at higher pressures it will increase significantly. Figure 1 illustrates this point for nitrogen, the principal component of air: the viscosity is 18 × 10-6 Pa.s up to 40 bar (atm) pressure, 20 × 10-6at 100 bar, and 53 × 10-6 at 1000 bar. The viscosities of air at several pressures from 1 to 100 bar are shown in Figure 2 as a function of absolute temperature. This shows that the effect of pressure increases at lower temperatures. Viscosities of a number of common gases at 1 bar are shown in Figure 3 to increase rapidly with absolute temperature, contrary to the behavior of liquids. The low viscosity of hydrogen is striking, as is the deviation of water vapor from the general trend. The water vapor curve terminates at its boiling point of 373 K. In determining viscosity as a function of temperature, two equations are often used. As can be seen from Figure 2, log (gas viscosity) is nearly linear with log (temperature) and can be represented by Equation 1. 0-8493-3904-9/97/$0.00+$.50 1997 by CRC Press LLC

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where µ0 is the viscosity at a known absolute temperature, T0(K). Exponent n is usually below one and is given in Table 1. The Sutherland formula fits the data slightly better than Equation 1 for hydrogen and helium:

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FIGURE 1 Viscosity of nitrogen at 300 K.

FIGURE 2 Viscosity of air at several pressures. For To = 300 K and T = 400 K, the constant C is 61 for hydrogen and 70.5 for helium based on the viscosity data in Table 1. The differences between Equations 1 and 2 are commonly so small that Equation 1 with the exponent n from Table 1 can normally be used. When other data are not available, viscosity of a gas may be estimated from its pressure and temperature at the critical point. The critical values for selected pure gases are given in Table 3, abstracted from Reference 1. In order to estimate the viscosity, first calculate the reduced temperature and pressure for the T and P for which the viscosity is desired. Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Viscosity of several gases at P = 1 bar.

These values may then be used in Figure 4 to estimate the reduced viscosity. To obtain the actual viscosity, one also needs the viscosity at the critical point. If the viscosity is not known at the critical point, it may be estimated from a known viscosity at some other Po and To. By calculating Pcr and Tcr from Equation 3, Figure 4 may be used to estimate µcp and then

The viscosity at the desired point is obtained by determining µr at the desired Pr and Tr from Figure 4, and then

If no viscosities are known, the critical viscosity may be estimated from the values of Pcr, Tcr, and molecular weight M, as follows:

Table 4 illustrates the application of the two methods to the calculation of the viscosity of nitrogen. For each temperature, Tr is calculated from Tcr = 126.3 K. Next are listed the values of Copyright © 1997 CRC Press, LLC.

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µr estimated from Figure 4 from the low density limit curve. Assuming we know the viscosity is 17.9 × 10-6 Pa s at 300 K where µr is estimated to be 1.00, the value of µcr is the same, and the viscosities at the other temperatures are directly calculated as shown. If no viscosities are known, Equation 6 is used:

Values for the other temperatures are calculated directly from the estimated values of µr The final column shows the actual viscosities, indicating a reasonable check. The following are conversions to the SI system: 1 reyn = 1.45 × 10-10 Pa.s and 1 P = 0.1 Pa.s.

SPECIFIC VOLUME The specific volumes listed in Table 1 indicate the degree of “perfection” of a gas. At 273.1 K and 1 bar pressure, 1 g-mol of a perfect gas occupies a volume of 22.4 1. Adjusting this to 300 K gives 24.6 1. If the specific volumes in Table 1 are multiplied by the molecular weight for each gas, the result is liters per gram mole (1/g-mol) and also cubic meters per kilogram-mol (m3/kgmol). Values for air, nitrogen, and oxygen are 24.9. Freon 21, Freon 11, and sulfur dioxide are below the perfect gas figure, indicating some degree of association between molecules. Pressure is given here in bars or atmospheres. For use in the SI system, 1 bar is equivalent to 101,300 Pa.

MEAN FREE PATH The mean free path is a measure of the average distance between collisions of the gas molecules. It is a function of the volume density of the gas and is given by Equation 7. Copyright © 1997 CRC Press, LLC.

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FIGURE 4 Generalized reduced viscosity of gases. (From Hougen, O. A., Watson, K. M., and Ragatz, R. A., C. P. P. Charts, 2nd ed., John Wiley & Sons, New York, 1960. With permission.)

where σ is the molecular diameter and n is the molecular density in molecules per cubic centimeter. The number of molecules per gram-mole is Avogadro’s number, 6.02 × 1023. As an example, the molecular density of argon at 273 K is

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The molecular diameter is approximately 2.9 × 10-8 cm. Applying Equation 7:

The value estimated in the Handbook of Chemistry and Physics is 9.0. This accuracy is quite sufficient for low pressure or very thin film bearing design.

SPEED OF SOUND Speed of sound in a gas is a function of temperature, molecular weight, heat capacity at constant pressure, and the gas constant:

Applying this data to oxygen, Table 1 lists M = 32.00 and Cp = 0.920 kJ/kg-K. Using this data in Equation 8 yields 329 m/s, as compared with the value of 353 m/s listed in Table 1.

NOMENCLATURE C Cp Cv P T TB U h M w v λ µ σ φ

= = = = = = = = = = = = = = =

Temperature, C Specific heat at constant pressure, kJ/kg⋅K Specific heat at constant volume, kJ/kg⋅K Pressure, N.m-2 Absolute temperature, K Boiling point, K Surface velocity, m/s Film thickness, m Molecular weight Width of leakage path, m Kinematic viscosity, m2/s Mean free path, m Absolute viscosity, N - s/m2= Pa.s Mass density, kg/m3 Sonic velocity, m/s

REFERENCES 1. Vargaftile, N. B., Tables on the Thermophysical Properties of Liquids and Gases, 2nd ed., John Wiley& Sons, New York, 1975. 2. Lide, D., Ed., Handbook of Chemistry and Physics, 75th ed., CRC Press, Boca Raton, FL, 1994, 6239. 3. Wilcock, D. F, Properties of Gases, in Handbook of Lubrication, Vol. 2, CRC Press, Boca Raton, FL,1984, 291-300. 4. Bird, R. B., Steward, W. E., and Lightfoot, E. N., Transport Phenomena, John Wiley & Sons, New York, 1960. 5. Uyehara, O. A. and Watson, K. M., Natl. Pet. News, 36, 764, 1944. 6. Hougen, O. A., Watson, K. M., and Ragatz, R. A., C. P. P. Charts, 2nd ed., John Wiley & Sons, New York, 1960. Copyright © 1997 CRC Press, LLC.

II Typical Lubrication Specifications for Equipment

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of Viscosity 17 Comparison Classifications E. Richard Booser

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FIGURE 1 Comparative viscosity classifications — for general guide only. Viscosities are based on a 95 VI oil. (From Engine Transmission and Axle Lubricant Classifications, Specifications, and Tests, Texaco Lubricants Co., Houston, TX, 1994. With permission.) Copyright © 1997 CRC Press, LLC.

18 Military Specifications Bobby D. McConnell CONTENTS Introduction.........................................................................................................................................................182 List of Military Specifications by Responsible Agency and Category in Numerical Order..........183 Tables......................................................................................................................................................................185

INTRODUCTION Military specifications are used by Department of Defense agencies to qualify and procure fluids and lubricants for military systems and equipment. These specifications, usually referred to as Mil Specs, are prepared and managed by the Army, Air Force, or Navy with coordination between the three services and DoD agencies such as DGSC, DFSC, DLA, etc. where necessary. A systematic numbering system is used with letters of the alphabet to designate the type of fluid or lubricant as well as the most recent version of the Mil Spec with dates of approval. For example, with MIL-L23699E, the L designates lubricant and the E means this the fifth and as of this writing the most recent version of this particular Mil Spec. Other letter designations include H for hydraulic fluids, G for greases, T for thread compound, etc., and are also described in the title of the Mil Spec. The letters are used for several different types of materials, i.e., C for coolant, cord, corrosion preventive, etc. The standardized format of Mil Specs includes: 1. SCOPE, which defines the use of the fluid or lubricant; 2. APPLICABLE DOCUMENTS, which lists references for other applicable Mil Specs, Standards, Technical Manuals, and nongovernmental publications such as ASTM Test Methods; 3. REQUIREMENTS, which describe the chemical, physical property, and performance requirements; 4. QUALITY ASSURANCE PROVISIONS, which describe inspection, quantification procedures, compliance, lot formation, sampling, test reports, and test methods; 5. PACKAGING, which defines preservation, packing, and marking; and, finally, 6. NOTES, which provide additional information for the manufacturer and/or user. A great deal of information is contained in the Mil Specs, such as the test methods which are used to determine the lubricant properties, but only the physical and chemical property requirements will be provided here. Most of the required properties are not absolute values but must meet a minimum or maximum value. Thus, there can be variations in the value of the property from one qualified source to another. If additional information is required, the reader should obtain the complete Mil Spec or call one of the agency managers. Contact the Mil Spec manager at one of the following responsible agencies: Air Force Lubricants Greases and Hydraulic fluids WL/MLSE Bldg. 652 2179 Twelfth St. Ste. 1 WPAFB, OH 45433-7718 937-255-7482

Air Force Turbine Engine Oils WL/POSL Bldg. 490 1790 Loop Road N WPAFB, OH 45433-7103 937-255-3100

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Army Oils, Greases and Solid Lubricants Navy Aircraft Oils and Greases AMSTA-RBF NAWS Aircraft Division 10101 Gridley Road Ste 104 Code 6061, PO Box 5152 Ft. Belvoir, VA 22060-5818 Warminster, PA 18974-0591 703-704-3722 215-441-1567 Navy Turbine Engine Oils NAWC Aircraft Division Code 445,PO Box 7176 Trenton,NJ 08628 609-538-6856 There is currently action within the DoD to consolidate and/or eliminate various Mil Specs where appropriate. Only those Mil Specs expected to remain active have been included in this handbook. Most Mil Specs have a Qualified Products List (QPL) which lists the products and manufacturers qualified to provide the fluids and lubricants under that Mil Spec. However, others are covered by 1st Article inspections, which means a sample of the material must be inspected before a given purchase is made. Since this information is constantly changing and cannot be included in the handbook, the reader is requested to contact the appropriate Mil Spec manager for this or any other information about a particular specification. This section includes a listing of the Mil Specs in numerical order, with title and category for each of the agencies, i.e., Air Force, Army, Navy, and a few DoD specs, concluding with a series of tables (in numerical order except for Table 1) of the physical and chemical property requirements for all the specs.

LIST OF MILITARY SPECIFICATIONS BY RESPONSIBLE AGENCY AND CATEGORY IN NUMERICAL ORDER Air Force HYDRAULIC FLUIDS MIL-H-5606G — Hydraulic Fluid, Petroleum Base, Aircraft, Missile, and Ordnance MIL-H-27601B — Hydraulic Fluid, Fire Resistant, Hydrogenated Polyalphaolefin Base, High Temperature, Flight Vehicle, Metric MIL-H-87257 — Hydraulic Fluid, Fire Resistant, Low Temperature Synthetic Hydrocarbon Base, Aircraft and Missile LUBRICANTS MIL-L-6081C — Lubricating Oil, Jet Engine MIL-L-6085C — Lubricating Oil, Instrument, Aircraft, Low Volatility MIL-L-6086D — Lubricating Oil, Gear, Petroleum Base MIL-L-7808K — Lubricating Oil, Aircraft Turbine Engine, Synthetic Base MIL-L-7870B — Lubricating Oil, General Purpose, Low Temperature MIL-L-87100 — Lubricating Oil, Aircraft Turbine Engine, Polyphenyl Ether Base MIL-L-87132B — Lubricant, Cetyl Alcohol, 1-Hexadecanol, Application to Fasteners GREASES MIL-G-27617E — Grease, Aircraft and Instrument, Fuel and Oxidizer Resistant MIL-G-83261A — Grease, Aircraft, Extreme Pressure, Antiwear MIL-G-83363B — Grease, Transmission, Helicopter MIL-G-83414 — Grease, Gunmount, Aircraft MISCELLANEOUS MIL-T-5544C — Thread Compound, Antiseize, Graphite-Petrolatum MIL-C-6529C — Corrosion Preventive, Aircraft Engine MIL-C-8188C — Corrosion Preventive, Oil, Gas Turbine Engine, Aircraft Synthetic Base Copyright © 1997 CRC Press, LLC.

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MIL-T-83483B — Thread Compound, Antiseize, Molybdenum Disulfide-Petrolatum MIL-C-87252B — Coolant Fluid, Hydrolytically Stable, Dielectric Army HYDRAULIC FLUIDS MIL-H-6083E — Hydraulic Fluid, Petroleum Base for Preservation and Operation MIL-H-46170B — Hydraulic Fluid, Rust Inhibited, Fire Resistant, Synthetic Hydrocarbon Base MIL-H-53119 — Hydraulic Fluid, Nonflammable, Chlorotrifluoroethylene Base LUBRICANTS MIL-L-2104F — Lubricating Oil, I-C Engine, Combat/Tactical Service MIL-L-2105D — Lubricating Oil, Gear, Multipurpose MIL-L-3150C — Lubricating Oil, Preservative MIL-L-14107C — Lubricating Oil, Weapons, Low Temperature MIL-L-21260D — Lubricating Oil, I-C Engine, Preservative and Break-In MIL-L-46000C — Lubricant, Semifluid (Automatic Weapons) MIL-L-46010B — Lubricant, Solid Film, Heat Cured, Corrosion Inhibiting MIL-L-46147A — Lubricant, Solid Film, Air Cured, Corrosion Inhibiting MIL-L-46150 — Lubricant, Weapons, Semifluid (High Load-Carrying Capacity) MIL-L-46167 — Lubricating Oil, I-C Engine, Arctic MIL-L-53131 — Lubricating Oil, Precision Rolling Element Bearing, Polyalphaolefin-Based MIL-L-63460D —Lubricant, Cleaner and Preservative for Weapons and Weapons Systems (Metric) GREASES MIL-G-10924F — Grease, Automotive and Artillery Navy HYDRAULIC FLUIDS MIL-H-22072C(AS) — Hydraulic Fluid, Catapult, NATO Code Number H-579 MIL-H-81019D — Hydraulic Fluid, Petroleum Base, Ultra-Low Temperature, (Metric) MIL-H-83282C — Hydraulic Fluid, Fire Resistant, Synthetic Hydrocarbon Base, Aircraft LUBRICANTS MIL-L-1970IB—Lubricant, All-Weather, Semifluid, for Aircraft Ordinance, Metric MIL-L-22851D — Lubricating Oil, Aircraft Piston Engine (Ashless Dispersant) MIL-L-23398D — Lubricant, Solid Film, Air-Cured, Corrosion Inhibiting MIL-L-23699E — Lubricating Oil, Aircraft Turbine Engine, Synthetic Base GREASES MIL-G-4343C — Grease, Pneumatic System MIL-G-6032D — Grease, Plug Valve, Gasoline and Oil Resistant MIL-G-21164D — Grease Molybdenum Disulfide, for Low and High Temperatures MIL-G-23549C — Grease, General Purpose MIL-G-23827B — Grease, Aircraft and Instrument, Gear and Actuator Screw MIL-G-25013E — Grease, Aircraft, Ball and Roller Bearing MIL-G-25537C — Grease, Aircraft, Helicopter Oscillating Bearing MIL-G-81322E — Grease, Aircraft, General Purpose, Wide Temperature Range MIL-G-81827A — Grease, Aircraft, High Load Capacity, Wide Temperature Range MIL-G-81937A — Grease, Instrument, Ultra-Clean, Metric MISCELLANEOUS MIL-M-7866C — Molybdenum Disulfide, Technical, Lubrication Grade MIL-S-8660C — Silicone Compound, NATO Code Number S-736 Department of Defense GREASES DoD-G-24508A(Navy) — Grease, High Temperature, Multipurpose (Metric) DoD-G-85733(AS) — Grease, High Temperature, Catapult System Copyright © 1997 CRC Press, LLC.

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LUBRICANTS DoD-L-25681D — Lubricant, Molybdenum, Silicone DoD-L-81846B — Lubricating Oil, Instrument, Ball Bearing, High Flash Point DoD-L-85336A — Lubricant, All-Weather (Automatic Weapons) DoD-L-85734(AS) — Lubricating Oil, Helicopter Transmission System, Synthetic Base

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19 Air Compressor Lubrication William Scales, P.E. CONTENTS Reciprocating Compressors.............................................................................................................242 Rotary Screw Compressors.............................................................................................................244 ISO Classification of Compressor Lubricants............................................................................244 Maintenance and Monitoring..........................................................................................................245 References...........................................................................................................................................247

A good air compressor or vacuum pump lubricant should minimize friction and wear, reduce internal leakage paths, protect against rust and corrosion, and leave virtually no deposits on hot discharge surfaces. Oil selection should be based on performance; the proper oil is probably neither the most expensive nor the least costly.

RECIPROCATING COMPRESSORS Reciprocating compressors fall into two general categories: 1. Single-acting compressors which compress air only in one direction of the piston stroke are also classified as automotive compressors or trunk piston units with cylinder lubrication supplied from the crankcase. These units normally range up to 100 hp, are generally aircooled, and are not intended for heavy duty applications at prolonged full load operation. While the function of the piston rings is to control and reduce oil carryover into the air stream, some oil will pass. One manufacturer has stated that one ounce of oil per 50 hp hours is acceptable. Petroleum-based lubricants, mineral oils with rust and oxidation inhibitors, are often recommended, as shown in Table 1, with viscosities varying between ISO 32 and ISO 68, dependent on operating conditions and designs. Unusually heavy or light duty applications and either cold or warm environments may require modifications to this very general specification. 2. Double acting compressors lubricate the frame or crankcase by constant circulation. This may require changing the oil once every 6 months. Generally, an external lubricator driven by a mechanical linkage off the crank shaft is adjusted to supply the cylinder with the proper amount of lubricant, which is measured unprecisely in drops. Valves and cylinders that are properly lubricated will have a light film of oil. Pools or concentrations in valve pockets indicate too much lubricant. Inadequate lubrication will be evidenced by dry cylinders and valves. Once the proper rate has been established, drops can be counted and used as the measurement. If the oil consumption rate changes, inspect the Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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lubricator for malfunctioning. If the compressor uses too much oil, it is important that the problem be corrected quickly. A formula for estimating the amount of oil to inject inot a cylinder is:

where B = bore (in.), S = stroke (in.), N = rpm, and Q = quarts of oil per 24-hour day. Most often, cylinder lubricants follow the same specifications as those used on single-acting compressors. Viscosities recommended by manufacturers vary according to piston and ring materials, cylinder sizes, and operating conditions: commonly an ISO 32 or 46 for smaller cylinders (up to 9inch diameter) and perhaps an ISO 100 or even 150 for cylinders 18 inches and larger. Reciprocating compressors have elevated compressed air discharge temperatures which range from 275 to 350°F (two-stage) to well above 500°F for single-stage models. A low carbon residue rating for oil is recommended, although this does not guarantee low carbon deposits in hot valve areas. The danger of fires and explosions is present in all lubricated compressors using petroleumbased lubricants. While infrequent, violent explosions can result where heavy carbon deposits have formed. Because of this, synthetic lubricants, especially diester based, have been used more extensively in the past few years. The primary objections have been high cost and possible incompatibility with materials in the compressed air system. If petroleum-base lubricants have been used, introduction of synthetics may dissolve the deposits and form a viscous tar, possibly causing damage to the intercoolers, aftercoolers, and Copyright © 1997 CRC Press, LLC.

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valves and even result in fire and explosion. Complete cleaning should be done before changing from petroleum-base to synthetic lubricant. In addition, be certain the synthetics are compatible with all materials used in the compressor and in all components downstream.

ROTARY SCREW COMPRESSORS In midsized manufacturing plants, the packaged rotary screw compressor with low initial cost and reduced maintenance has become the choice for most installations up to 1500 cfm. These compressors are available from 20 cfm (5 hp) up to 3000 cfm (600 hp) at 100 psig. Discharge air temperatures are normally maintained below 200°F to minimize oxidation, but oil injection into the compression chamber for removing the heat of compression is usually about 145 to 150°F to avoid condensation of water vapor from the compressed air. Some manufacturers still recommend automotive engine oils meeting API specifications such as CE/SK or MIL L-2104 B,C,D; others specify turbine oil or transmission fluid. During recent years, however, most manufacturers of rotary screw air compressors have contracted with lubricant suppliers to sell brand-label synthetic lubricants exclusively for their compressors. Because of the wide variance in requirements and opinions among manufacturers, no recommendations can be given that will cover all rotary compressors. Since manufacturers change their specifications and new oils are being developed, call the manufacturer to get the most recent recommendations. Most manufacturers have selected PAO synthesized hydrocarbons with a specific additive package for rotary compressors. Other manufacturers have specified diester fluids, polyglycols, or semisynthetics. One advantage of synthetics in rotary compressors is extended life, often three to four times that of standard petroleum-based lubricants. Some suppliers have received FDA authorization to use their lubricants in compressors where the compressed air may have incidental contact with food. However, limitations of the lubricant additive package reduces the life expectancy to about half that of normal synthetics.

ISO CLASSIFICATION OF COMPRESSOR LUBRICANTS ISO 6743 establishes the detailed classification of lubricants for use in air compressors, gas compressors, and refrigeration compressors. ISO 6743-3 (in conjunction with ISO 6743-0) provides a rationalized range of the most common internationally available compressor lubricants without restrictions by specifications or product descriptions. Detailed classification of each category of compressor lubricants in Tables 2 and 3 involves a group of three letters. The first letter, “D,” identifies a compressor lubricant, and the second and third letters designate the specific type. These letters are then followed by the viscosity grade according to ISO 3448. For example, a severe-duty air compressor oil may be designated ISO-L-DAB 68, or abbreviated L-DAB 68.

ISO REFERENCE STANDARDS: ISO 3448 Industrial liquid lubricants — ISO viscosity, classification. ISO 5388 Stationary air compressors — Safety rules and code of practice ISO 6521 Lubricants, industrial oils and related products (class L) — Specifications of categories L-DAA and L-DAB (Lubricants for reciprocating and drip-feed rotary air compressors). ISO 6743-0 Lubricants, industrial oils and related products (class L) — Classification Part 0: General. ISO 8681 Petroleum products and lubricants — Method of classification — Definition of classes.

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MAINTENANCE AND MONITORING Important: 1. NEVER MIX LUBRICANTS, or change from one type to another, unless complete draining of the old oil and a thorough cleaning precede the changeover. This applies to petroleum base and synthetics equally. 2. Obtain the air compressor manufacturer’s approval before changing lubricants. In some cases, the lubricant may not be compatible with materials used in the compressor. There have not been standards specifically established for synthetic lubricants, but many of the same physical properties, characteristics, and additives are required as listed in Table 1. While synthetics are often sold on the basis of having a life expectancy of four times that of petroleumbased oils, it is extremely important to maintain an analysis program which samples the lubricant at intervals recommended by the lubricant supplier. The sampling program may provide basic information such as antioxidant levels, particle counts, total acid number (TAN), viscosity, and

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spectrochemical anylysis, listing the presence of various metals and water in the sample. Even with such a monitoring program, however, the printed life expectancy of the lubricant should not be exceeded unless approved by the supplier because depletion of corrosion inhibitors, rust preventatives, and other important additives may not be reported.

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REFERENCES 1. W. Scales, Air Compressors and the Compressed Air System, Complete Building Equipment Maintenance Desk Book, 2nd ed., Prentice Hall, Englewood Cliffs, NJ, 1992. 2. Compressed Air and Gas Handbook, 5th ed., Compressed Air and Gas Institute, Cleveland, Ohio, 1989. 3. Compressor Engineering Data, Scales Air Compressor Corporation, Carle Place, New York, 1992.

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20 Automatic Transmission Fluids* Ron Desing Automatic transmission fluids (ATFs) are among the most complex multifunctional fluids available because they serve many purposes in the transmission. In general, ATF performance is defined by the service-fill specifications of passenger-car and commercial-vehicle transmission manufacturers. These specifications establish both testing procedures and pass/fail criteria for the performance parameters listed above. Measuring the frictional characteristics of the fluid in contact with the transmission’s frictional elements is especially critical in ensuring acceptable driveability for a particular vehicle. Since January 1, 1994, General Motors has required the use of fluids certified as meeting its DEXRON-III** specification in its passenger and commercial vehicles equipped with automatic transmissions. Fluids that have not been formally approved against these requirements should not use the DEXRON term. These fluids are “back serviceable” or suitable for older GM transmissions. Each DEXRON-III fluid marketed has a unique “F-number” designation. Ford Motor Co. has required the use of fluids meeting the MERCON*** specification since 1987. The specification has undergone significant revisions since its original release. Since January 1, 1994, only fluids meeting the latest requirements may display the trademark term. These formulations are identified in the market by an identifying code (“M-number”) beginning M 93 or higher. MERCON fluids also are “back serviceable” to Ford vehicles produced since 1981. Generally, vehicles produced prior to that model year require a different type of fluid, Ford Type F. The use of Type F fluids in transmissions designed for MERCON, or vice versa, could result in poor shift quality. The Allison Transmission Division of General Motors specifies fluids meeting the requirements of the Allison C-4 specification for their heavy-duty transmissions used in commercial and off-highway vehicles. Caterpillar now requires the use of straight grade fluids meeting its TO-4 specification for its transmissions, wet-disc brakes, and final drives. Other transmission makers have different ways to identify suitable fluids. European original equipment manufacturers (OEMs) such as Mercedes-Benz and transmission manufacturers Zahnradfabrik Friedrichshafen (ZF), Voith, and Renk do not have a service-fill system, per se. Instead, they rely on lists of lubricants that they have approved for use. Similarly, Japanese OEMs do not use a service-fill specification and recommend only their part number fluids, or “genuine” oils, for service work. While ATF performance is defined by the specifications of automotive OEMs, a significant percentage of the fluids produced is used in applications other than automatic transmissions. These include: * From Ready Reference for Lubricant and Fuel Performance, The Lubrizol Corp., Wickiffe, OH, 1995, pp. 113-121. With permission.. ** DEXRON is a registered trademark of General Motors Corp. *** MERCON is a registered trademark of Ford Motor Co.

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• Powershift transmissions in off-highway construction, agricultural and mining equipment • Automotive, industrial, mobile, and marine hydraulic systems • Power steering units • Rotary screw compressors Automatic transmission fluids typically contain antioxidants, antifoam agents, viscosity modifiers, antiwear additives, friction modifiers, and seal swell modifiers. They are dyed red for identification purposes.

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Automotive Engine Lubricants Yeau-Ren Jeng and Ron Desing

CONTENTS Engine Lubricant Constituents....................................................................................................253 Physical Requirements for Engine Oils......................................................................................253 Engine Oil Classification and Requirements.............................................................................253 Abbreviations: Organizations Referenced in Tables................................................................255

ENGINE LUBRICANT CONSTITUENTS LUBRICANT BASE STOCKS: Mineral oil Synthetic oil

Refined from crude oils Derived from chemical reactions Synthetic base stocks Olefin oligomers (PAOs) Dibasic acid and polyol esters Alkylated aromatics Polyaklylene glycols Phosphate esters

Mineral base stocks Paraffins Naphthenes Aromatics

PHYSICAL REQUIREMENTS FOR ENGINE OILS SAE has established that twelve viscosity grades are suitable for engine lubricating oils. The physical requirements for these viscosity grades are described in SAE J300, which is intended for use by engine manufacturers in determining engine oil viscosity grades suitable for use in their engines. Current viscosity requirements for these grades are shown in Table 1. The U.S. military imposes additional requirements for the grades in Table 2.

ENGINE OIL CLASSIFICATION AND REQUIREMENTS The classification of and requirements for engine oils are set by a tripartite arrangement (SAE, API, ASTM) in the U.S. In Europe, engine oil classifications represent a mixture of a quadripartite agreement (ACEA, CEC, ATC, ATIEL) and individual OEM requirements. In the U.S. API has standardized the labeling of engine oils by adopting the so-called donut logo. The top half, the “API Service” section, indicates the oil quality level; the lower half, the Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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“Energy Conserving” section, states the oil’s degree of fuel efficiency; if fuel economy claims are not made the low section of the logo is left blank; and the center circle represents the viscosity level.

OIL PERFORMANCE RATINGS Oils are rated by how they perform in engine tests, rather than on what they contain. These tests, developed jointly by the auto and oil industries and allied technical societies, represent different engine requirements. Thus, an oil may be rated only for vehicles up to a certain model year, or it may be rated for use only in gasoline engines. The various ratings are labeled by the API and are called “service categories.”

API SERVICE CATEGORIES FOR GASOLINE ENGINES SH is the highest quality oil available for use in previous and current models. Engine test requirements are CRC L-38, Sequence IID, Sequence HIE, and Sequence VE. Category SH was adopted in 1992 to describe engine oil first mandated in 1993. It is for use in service typical of gasoline engines in present and earlier passenger cars, vans, and light trucks operating under vehicle manufacturers’ recommended maintenance procedures. Engine oils developed for this category provide performance exceeding the minimum requirements of API Service Category SG, which it is intended to replace, in the areas of deposit control, oil oxidation, wear, rust, and corrosion. Oils meeting API SH requirements have been tested according to the Chemical Manufacturers Association (CMA) Product Approval Code of Practice and may utilize the API Base Oil Interchange and Viscosity Grade Engine Testing Guidelines. They may be used where API Service Category SG and earlier categories are recommended. Copyright © 1997 CRC Press, LLC.

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ABBREVIATIONS: ORGANIZATIONS REFERENCED IN TABLES ACEA

Association des Constructeurs Europeens de l’ Automobile (Association of European Automotive Manufacturers) ASTM American Society for Testing and Materials ATC Technical Committee of Petroleum Additive Manufacturers (Europe) ATIEL Association Technique de l’ lndustries Européenne des Lubrifiants Copyright © 1997 CRC Press, LLC.

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API American Petroleum Institute CCMC Comité des Constructeurs d’Automobiles du Marche Commun (replaced by ACEA) CEC Couseil Europeen de Coordination pour les Developments des Essais de Performance des Lubrifiants et des Combustibles pour Moteurs (Coordinating European Council) CRC Coordinating Research Council (U.S.) DKA Deutscher Koordinierungsausschuss im Coordinating European Council (member CEC) ILSAC International Lubricant Standardization and Approval Committee JASO Japan Automobile Standards Organization JIS Japanese Industrial Standards SAE Society of Automotive Engineers

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Diesel Engine Lubricants* Ron Desing

CONTENTS Status of API “C”.........................................................................................................................263 Tables...............................................................................................................................................265 Abbreviations..................................................................................................................................271

STATUS OF API “C” COMMERCIAL ENGINE SERVICE CATEGORIES (FOR FLEETS, CONTRACTORS, FARMERS, ETC.)

CF-4 — 1990 DIESEL ENGINE SERVICE Service typical of high-speed, four-stroke cycle diesel engines. API CF-4 oils exceed the requirements for the API CE category, providing improved control of oil consumption and piston deposits. These oils should be used in place of API CE oils. They are particularly suited for onhighway, heavy-duty truck applications. When combined with the appropriate “S” category, they can also be used in gasoline and diesel-powered personal vehicles — i.e., passenger cars, light trucks, and vans — when recommended by the vehicle or engine manufacturer. CF — INDIRECT-INJECTED DIESEL ENGINE SERVICE Serivice typical of indirect-injected diesel engines that use a broad range of fuel types, including those using fuel with high sulfur content; for example, over 0.5% wt. Effective control of piston deposits, wear, and copper-containing bearing corrosion is essential for these engines, which may be naturally aspirated, turbocharged, or supercharged. Oils designated

* From Ready Reference for Lubricant and Fuel Performance, The Lubrizol Corp., Wickliffe, OH, 1995, pp. 53-71. With permission.

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for this service have been in existence since 1994 and may be used when API Service Category CD is recommended. CF-2 — SEVERE-DUTY TWO-STROKE CYCLE DIESEL ENGINE SERVICE Service typical of two-stroke cycle diesel engines requiring highly effective control over cylinder and ring-face scuffing and deposits. Oils designed for this service have been in existence since 1994 and may also be used when API Engine Service Category CD-II is recommended. These oils do not necessarily meet the requirements of API CF or CF-4 unless they pass the test requirements for these categories. CG-4 — 1994 SEVERE-DUTY DIESEL ENGINE SERVICE API Service Category CG-4 describes oils for use in high-speed four-stroke-cycle diesel engines used in both heavy-duty on-highway (0.05% wt sulfur fuel) and off-highway (less than 0.5% wt sulfur fuel) applications. CG-4 oils provide effective control over high-temperature piston deposits, wear, corrosion, foaming, oxidation stability, and soot accumulation. These oils are especially effective in engines designed to meet 1994 exhaust emission standards and may also be used in engines requiring API Service Categories CD, CE, and CF-4. Oils designed for this service have been in existence since 1994. In addition to engine tests, diesel engine oils are sometimes required to show adequate performance in wet brake and transmission applications. Such testing is conducted in the Caterpillar TO-4 and Allison C-4 tests.

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ABBREVIATIONS: ORGANIZATIONS REFERENCED IN PRECEDING TABLES ACEA Association des Constructeurs Européens de 1’Automobile (Association of European Automotive Manufacturers) API American Petroleum Institute ASTM American Society for Testing Materials CCMC Comité des Constructeurs d’Automobiles du Marché Commun (replaced by ACEA) CEC Conseil Européen de Coordination pour les Developments des Essais de Performance des Lubrifiants et des Combustibles pour Moteurs (Coordinating European Council) CRC Coordinating Research Council (U.S.)

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23 Electric Motors and Generators E. R. Booser

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Hydraulic and Turbine Oil Specifications Ian Macpherson and Daniel D. McCoy

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Hydraulic and Turbine Oil Specifications

REFERENCES 1. Lubricant Specification Handbook, Ethyl Corp., Richmond VA, 1990, 1991, 1994. 2. Ready Reference for Lubricant and Fuel Performance, Lubrizol Corp., Wickliffe, OH, 1993.

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25 Marine Diesel Engines* The performance of lubricants for marine engines is not defined by a standard classification system. Standard dynamometer test methods or testing protocols leading to industry-wide approvals are not available. Performance levels and the approval process are driven by the major engine builders, most of whom publish list of lubricants approved for use in their engines. They encourage their customers to use lubricants appearing on the approval list. In general, OEMs, a full-scale ship trail lasting 5000 hours (about 1 year) to approve a lubricant. Engine builders approve only formulation specific lubricants; blanket approvals for additives are not available. Marine engines generally use residual fuels with high sulfur content (2 to 5% by mass), although fuels can vary widely around the world. Because the cost of fuel is a significant portion of a ship’s total operating cost, engine designers are optimizing their engines to cut fuel consumption, and shipowners purchase the lowest cost fuels available. These two factors are placing greater demands on the quality of the oils used for marine engine lubrication. Two types of engines dominate propulsion on large ocean-going vessels. Parameters for each type are listed below:

Two-stroke slow-speed crosshead engines requires two lubricants: one for the upper cylinder (cylinder oil) and one for the crankcase (system oil). MAN B&W and Sulzer dominate the market for these engines. (Together they account for 90% of the market.) Their requirements for cylinder oils are generally SAE 50 viscosity grade and 70 TBN. System oils are usually SAE 30 and 5 to 10 TBN. Four-stroke medium-speed trunk piston engines require only on lubricant because they have a common sump for the crankcase and cylinder. The market for these engines is more fragmented than that for two-stroke engines, and several builders hold a significant position. Wartsila, Pielstick, Sulzer, MAN B&W, and Mak are among the largest OEMs in this market. Lubricants for medium-speed engines are generally SAE 30 or 40 viscosity grade; TBN requirements vary from 12 to 40, depending on fuel sulfur content.

* From Ready Reference for Lubricant and Fuel Performance, Lubrizol Corp., Wickliffe, OH, pp. 107-108, 1995. With permission.

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26 Railroad Diesel Engines* There are no generally accepted tests for assessing railroad oil quality. Each engine builder has its own preferences based on field experience, and only quality in an extended road test can earn approval for a product. Most commercial oils are compounded with additives to meet the API CD classification, and they containl a generous reserve of alkalinity to compensate for the high sulfur content of the fuels burned. Some compositional restraints are imposed — for example, zinc limits — to prevent engine component damage. Most oils covered in Table 1 are SAE 40 single grades, but use of multigrade formulations is expanding to improve fuel economy.

In the U.S., the Locomotive Maintenance Officers Association Fuels and Lubricants Committee has established the Generation designation in Table 2 for railroad diesel engine oils. The highest quality oils are designated “Generation 5” and must pass the OEM evaluation tests in Table 3 and meet the following requirements:

* From Ready Reference for Lubricant and Fuel Performance, Lubrizol Corp., Wickliffe, OH, 1995, pp, 105-106. With permission.

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• Allow drains of 180 days minimum for low-oil-consuming engines averaging 10,000 miles

per month or consuming fuel (0.3 to 0.5% sulfur) at a rate of 20,000 gallons per month. • Pass OEM oxidation, corrosion, and friction tests. • Meet OEM engine test requirements. • Pass OEM field test requirements.

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Cincinnati Milacron Slideway Oil Specifications Ian Macpherson

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III Industrial Application Practices

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28 Aluminum Production Equipment Edward J. Myers CONTENTS Key to Lubicants for Aluminum Production Equipment...........................................................289 Antifriction Bearing Oil Lubrication................................................................................................290 Grease Lubrication of Antifriction Bearings..................................................................................290 Electric Motor Bearing Grease Lubrication...................................................................................291 Electric Motor Bearing Oil Application..........................................................................................291 Journal Bearing Oil Lubrication........................................................................................................291 Power Train Lubrication.....................................................................................................................292 Enclosed Chain Drives, Gear Head Motors, PIV Drives, Etc....................................................293 Hydraulic System Oil Fluid Practices...............................................................................................293 Hydraulic Fluid Maintenance.............................................................................................................293 Way Lubrication....................................................................................................................................294 Air Compressor Lubrication..............................................................................................................294 Mobile Equipment...............................................................................................................................296 Standard Practice for Acceptance of Lubricants...........................................................................296 Storage and Dispensing.......................................................................................................................297

KEY TO LUBRICANTS FOR ALUMINUM PRODUCTION EQUIPMENT CO GO HO MO WO ATF EPG FR

Low deposit compressor oil. Extreme pressure gear oil. Rust and oxidation inhibited industrial and hydraulic oil. Internal combustion oil — followed by API service classifcation. Anti-stick–slip oils for machine tool slide ways and related hydraulic service. Automotive automatic transmission fluid. Extreme pressure general purpose lithium grease. Fire resistant hydraulic oil.

Viscosity of oils and grease grade indicated by these suffixes in parenthesis: SAE VG AGMA NLGI

Number for automotive oils. Viscosity grade (centistokes at 40(C) for industrial oils. Number for gear oil viscosity grade. Penetration grade for grease consistency.

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ANTIFRICTION BEARING OIL LUBRICATION

GREASE LUBRICATION OF ANTIFRICTION BEARINGS Grease relubrication frequency based on annual hours of operation at maximum bearing temperature of 65°C (150°F), with lubricant type EPG (NLGI 2).

Reduce above frequency 50% for each 15°C (25°F) bearing temperature increase above 65°C (150°C) or where bearings are subjected to solvents, vibration, chemicals, or hostile environment. Bearing operating temperature should be limited to 80°C (175°C) where possible to reduce lubricant oxidation. For bearing application above 121°C (250°F), only heat-stabilized bearing should be used and special grease may be required. Bearings should be provided with both grease fittings and drain relief fittings. Shots of grease from a medium (7,000 psi) grease gun to relubricate antifriction and plain bearings.

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ELECTRIC MOTOR BEARING GREASE LUBRICATION Relubricate with EPG (NLGI 2) grease for temperatures between 0°C (32°F) and 121°C (250°F).

The above frequency should be reduced by 50% for each 15°C (25°F) bearing temperature increase above 65°C (150°F) or where the motor is subjected to solvents, vibration, chemicals, abrasives or temperature.

ELECTRIC MOTOR BEARING OIL APPLICATION Lubricant HO (VG – 68) Check oil level on large MG sets daily and add oil as required. Where bearings are checked each shift by an electrician, lubrication personnel need only check and top off systems on a weekly basis. Change oil annually or every 6000 hours of operation. If the bearing oil system is suspected of bearing contamination, the unit should be shut down and the bearing and/or the bearing oil system drained, flushed, and refilled with new oil. An excess usage of oil should be investigated as oil contamination of the insulation will greatly reduce the expected life of the unit.

JOURNAL BEARING OIL LUBRICATION Automatic lubrication should be considered for any system requiring more than weekly maintenance checks. All circulating systems should be checked annually for contamination. Systems not checked should be drained and refilled annually. If suspected of being contaminated, the system should be shut down and bearing oil system drained, flushed, and refilled with new oil.

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ENCLOSED CHAIN DRIVES, GEAR HEAD MOTORS, PIV DRIVES, ETC.

HYDRAULIC SYSTEM OIL FUND PRACTICES All new hydraulic systems should be designed with “L” type reservoirs (reservoir along side pump) and flooded pump inlet. Pumps should be capable of operating on high water-base fluids or 90/10 emulsions.

Hydraulic system design should be limited to 1,000 psi except for rolling mill roll jacks, hydraulic mill cylinders, and extrusion presses. When filling or adding make-up fluid to any hydraulic system, all fluid will be filtered through at least to 10 mm filter element equipped with a maximum number of magnets.

HYDRAULIC FLUID MAINTENANCE Hydraulic fluids used in low-pressure systems operating below 2000 psi should be analyzed at least once a year to determine the following: Approximate viscosity by checking with a Visgauge. Allowable range ±15% of mean viscosity. Any variation greater should be suspected of being the wrong oil for the application. Pentane insolubles should be no more than 0.02% by volume and contain no more than a trace of iron.

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Hydraulic oil used in high-pressure systems on extrusion presses equipped with Oilgear pumps should be checked quarterly for contamination. Particle contamination should be maintained below 800 particles (>10 µm/ml) or 80,000 particles (>10 µm/100 ml) sample. Systems utilizing more that 50% makeup as reclaimed oil should be checked semiannually for viscosity, rust test, flash point, and rotary bomb oxidation. These procedures should improve pump life, extend oil life, lower total maintenance cost, and give advance warning of impending pump problems.

WAY LUBRICATION

AIR COMPRESSOR LUBRICATION Piston Compressor Lubricant CO (VG 68, 100, and 150)

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Rule of thumb: 0.1 pints per 500,000 ft2 of swept cylinder wall area. Experience indicates: Double-acting compressors: Small = 0.02 pints per 60,000 ft.2 of swept cylinder wall area/hour; large = 0.01 pints per 60,000 ft.2 of swept cylinder wall area/hour. During periods of initial break-in, the recommended feed rate should be doubled. Periodic inspection of the cylinder bore should be made. If the walls of the cylinders are dry, increase oil feed. If there is excessive oil in the cylinder and discharge line, reduce the oil feed. High oil feed can cause excessive carbon on the valve with loss of efficiency and could cause a fire and/or explosion in the air supply system. Annual maintenance should include inspection of the valves, drain and refilling of the crank case, inspection of intercoolers, aftercoolers, and air receivers. Compressors should never be run without the intercooler or aftercooler operating properly. Table below lists the theoretical adiabatic air temperature at compression pressures for single and two-stage units.

Rotary Screw Compressors. The following are minimum guidelines for rotary screw compressor lubrication and maintenance. In all cases, the manufacturer’s recommendations will be reviewed and take precedence.

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MOBILE EQUIPMENT (NOT INCLUDING MINING, ORE HANDLING, OR CONSTRUCTION)

STANDARD PRACTICE FOR ACCEPTANCE OF LUBRICANTS All bulk deliveries by truck transport or tank car must be accompanied by a certified analysis and must be delivered to the responsible plant personnel before unloading for review and acceptance. All drums shall be marked on the opening end with the supplier brand name and refinery lot number. Field Ispection—All material will be visibly inspected and any material in question will be set aside for inspection by the supplier’s representative. Each lot of lubricant will be checked for Copyright © 1997 CRC Press, LLC.

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viscosity by using a Model No. 38 Visgauge. Any variation of more than 10% of the mean viscosity should be questioned. For oils with a viscosity greater than 200 cSt, the difference may vary 20% and will only give comparative readings.

STORAGE AND DISPENSING All drums will be stored inside or in racks providing for horizontal storage of drums. Racks will be constructed to allow for inventory rotation. All dispensing areas will be covered areas of a building, free of excessive dust. Racks will be provided for dispensing lubricants. Pumps will be provided for dispensing grease, oil, and other lubricants. All drums are to have closed tops. Any open drum is to be discarded.

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Farm and Off-Highway Machinery*

Farm tractors, off-highway machinery, industrial tractors and related equipment require a variety of lubricants, ranging from engine oil to hydraulic fluids and transmissions oils to wet brake fluids. Except for engine oils, worldwide performance standards for tractor lubricants are sent by original equipment manufacturers. The most common standards are summarized in Table 1 to 4. Today’s tractor fluids lubricate transmissions, final drives, we brakes and clutches, and hydraulic systems, typically from a common fluid reservoir on the tractor. These unique performance characteristics make tractor fluids acceptable for use in both on and off-highway commercial transmissions and high-pressure hydraulic systems. They can also be used in mining equipment. Because of this versatility, these fluids are often called universal tractor transmission oils (UTTOs) in North America. In Europe and some other areas of the world, farmers have accepted the concept of super tractor oil universal (STOU). STOU oils are formulated to be used in the engine, wet brakes, hydraulic systems, and transmissions of many farm tractors currently in service.

* From Ready Reference for Lubricant and Fuel Performance, Lubrizol Corp., Wickliffe, OH, 1995, 97-104. With Permission.

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30 Food Industry Lubrication George R. Arbocus CONTENTS Introduction..........................................................................................................................................303 Maintenance Guidelines......................................................................................................................304 Typical Lubricant Applications in the Food Industry...................................................................304 References..............................................................................................................................................308

INTRODUCTION Processing of food and beverages requires special care in lubricant selection and application to meet the unique sanitation and toxicity needs of the industry. To meet these requirements, the federal government exercises direct control over food grade lubricants through standards issued by two agencies: the U.S. Department of Agriculture (USDA) regulates primarily meat and poultry plants, while the Food and Drug Administration (FDA) monitors other food manufacture as well as drug production.1,2 Upon satisfactory evaluation of nontoxicity, USDA issues one of two ratings: H1 for use as a lubricant where there is incidental contact as by splashing or dripping from machinery above an edible product, or H2 for use where there is no contact, as in sealed gear boxes or machinery below an edible product line. These classes include a number of petroleum and synthetic oils and greases. For severe requirements where lubricants contact food on a regular basis, FDA published a list of authorized ingredients in the Code of Federal Regulations.3 The following three classes of lubricants are identified: 1. White mineral oil (21 CFR 172.878) used, for example, as a release agent for bakery products, confections, dehydrated fruits and vegetables, and egg whites. 2. Petrolatums (21 CFR 172.880), used in applications similar to white mineral oils. 3. Technical white oils (21 CFR 178.3620), used in processing aluminum foil for food packaging, in manufacture of animal feed and fiber bags, and on food machinery.

These are highly purified petroleum products which have been fully refined by either acid treatment or hydrogenation to remove all unsaturates, aromatics, and coloring materials to meet U.S. Pharmacopoeia (USP) requirements. While the lubricant applications that follow are based on previous use, equipment manufacturers should be consulted on specific cases. Much of the equipment utilized in food plants is of unique design, often operates in wet or moist environments, and may be exposed to high pressure hoses and cleaners for maintaining the required degree of sanitation. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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MAINTENANCE GUIDELINES Lubricant selection • Select lubricants per machine manufacturer’s instructions. Lubricant storage • Store indoors in clean, dry areas. • Avoid contact with hot pipes, ovens, etc. • Use older stock first. Lubricants in use

• Transfer containers must be clean. • Do not use glass or galvanized metal transfer containers, use plastic. • Avoid contamination! Replace lids and bungs. Grease applications

• Flush old grease out when changing brands. • Isolate guns used for “food grade” greases. • Remove vent plug before applying, replace plug after new grease appears. • Over-greasing can damage seals and contaminate product. Oil applications

• Check level when equipment is idle. • Foaming can be due to incorrect oil level, moisture, air, leaks, or wrong oil. • In wet applications, allow water to separate, drain off, replace with fresh oil to proper level. Bearing care

• Keep bearings in factory container — wrapped until ready to use. • Never use compressed air to dry or spin bearings. • Check housing and shaft fits before mounting. Check print or bearing manufacturer for corect dimensions. Hydraulic systems

• Keep oil clean! • Replace or service oil filters as recommended. • Clean strainers regularly (monthly or as required). • Don’t change oil because it looks old — analyze! Air compressors • Check oil level daily. • Service air intake filters monthly, more often in dusty environment. • Service or change oil filters per manufacturer’s recommendations.

TYPICAL LUBRICANT APPLICATIONS IN THE FOOD INDUSTRY Note: Refer to manufacturer’s recommendations to verify lubricant type and grade.

BAKERIES Dough Mixer, Sifter, Divider General greasing: NLGI #2 USDA “H-l” aluminum complex grease Gear drives: ISO 320 or ISO 680 gear oil (check manufacturer) Open gears: over product line: USDA “H-l” tacky grease; below product line — industrial compound Chains and general: ISO 150 or ISO 220 USDA “H-l” products where contact is possible Oil applications: ISO 150 or ISO 220 where no product contact anticipated

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Proofers and Coolers Chains, product contact possible: ISO 150 “H-1” oil; no contact anticipated: ISO 150 tacky oil Gear drives: ISO 320 or ISO 680 gear oil Bearings: NLGI #2 “H-l” aluminum complex grease

Ovens Bearings in ovens: Synthetic high temperature NLGI #2 grease Bearings outside of ovens: NLGI #2 aluminum complex grease Oven chains: graphite or molybdenum disulfide oil General oiling: ISO 150 lubricating oil BREWERIES

Brew House Beer cocks, lauter tubs (stuffing box), oil plungers, miscellaneous oil applications: ISO 68 “H-l” General grease Product contact possible: NLGI #2 “H-l” grease No product contact: NLGI #2 lithium or aluminum complex

De Palletizers Airline oil: ISO 32 oil Hydraulic system: ISO 32 or 46 oil Bearings: NLGI #2 lithium or aluminum complex grease Gear drives: ISO 220 or 460 gear oil

Packaging Fillers — valves Grease: NLGI #1 or 2 “H-l” grease Oil: ISO 100 “H-l” oil Can closers Grease: NLGI #1 or 2 “H-l” grease Oil: ISO 100 “H-l” oil

Pasteurizers Chain: tacky ISO 150 chain oil Gear drives: ISO 460 or 680 gear oil Grease: NLGI #2 aluminum complex

Palletizers Airline oil: ISO 32 oil Hydraulic system: ISO 32 or 46 oil Bearings: NLGI #2 aluminum complex grease Gear drives: ISO 460 gear oil

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AIRIES

Milk Receiving Truck pusher, dock leveler: ISO 32 or 5W20 hydraulic oil Conveyors, rails: NLGI #2 aluminum complex grease Gear drives: ISO 220 gear oil

Compactor Hydraulic power unit: ISO 32 or 5W20 hydraulic oil

Milk Processing Mixing tank gear drives: ISO 460 gear oil Homogenizer gear drive: ISO 460 gear oil

Fillers (Various Types) Grease fittings: NLGI #2 grease Gear drives: ISO 220 or 460 gear oil Airline lubricator: ISO 32 hydraulic oil Chain drives: tacky ISO 150 chain lubricant

Automatic Casers Grease fittings: NLGI #2 grease Airline lubricator: ISO 32 hydraulic oil

Case Conveyors Chain drives: tacky ISO 150 chain oil Gear drives: ISO 220 gear oil Airline lubrication: ISO 32 hydraulic oil Grease fittings: NLGI #2 grease ICE CREAM MANUFACTURE Note: Some dairies may operate special equipment to prepare ice cream products. This equipment frequently requires lubricants similar to those used in milk processing operations. Lubricants with low-temperature performance are also utilized. FRUIT AND VEGETABLE PROCESSING

Bearings Grease, normal conditions Possible product contact: NLGI #2 “H-l” grease No contact anticipated: NLGI #2 aluminum complex or lithium base grease Grease for acids and alkali conditions: NLGI #2 chemical resistant grease Oil lubrication, normal conditions: ISO 100 or 150 rust and oxidation inhibited Oil lubrication, acid, and alkali conditions: consult lubricant supplier

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Fillers Valves: NLGI #1 or #2 “H-l” aluminum complex grease Gear drives: ISO 220 or ISO 460 gear oil Chains: ISO 150 “H-l” tacky lubricating oil

Can Closers Feed chains: “H-l” ISO 150 lubricating oil Bearings (hand gun): NLGI #2 “H-l” grease or NLGI #2 aluminum complex or lithium grease Bearings (lubrication system): NLGI #0 or #1 “H-l” aluminum complex grease or NLGI #0 or #1 lithium grease Gear drives: ISO 220 or 460 gear oil Can covers in stack: semifluid “H-l” lubricant

Freezing/Cold Storage Bearings, hand gun: low-temperature NLGI #2 lithium grease Refrigeration compressors Ammonia service: ISO 68 naphthenic oil Other refrigerants: check manufacturer’s recommendation SOFT DRINK PLANTS

De Palletizers Airline oil: ISO 32 lubricating oil Hydraulic system: ISO 32 or 46 hydraulic oil Bearings: NLGI #2 lithium or aluminum complex grease Gear drives: Check manufacturer’s recommendation

Fillers, Crowners Bearings (hand gun): “H-l” NLGI #1 or #2 grease Gear drives: ISO 220 or 460 gear oil Chain drives: “H-l” ISO 100 lubricating oil

Can Closers Bearings (hand gun): NLGI #2 “H-l” grease or NLGI #2 bearing grease Bearings (lubrication system): NLGI #0 or #1 “H-l” grease or NLGI #0 or #1 grease Gear drives: ISO 220 or 460 gear oil Chains: “H-l” ISO 100 lubricating oil Can covers in stack: semifluid “H-l” lubricant

Labelers Grease gun: “H-l” NLGI #2 aluminum complex grease or NLGI #2 bearing grease Open gears, slides, and cams: tacky “H-l” grease or industrial open gear lubricant aerosol

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REFERENCES AND ADDITIONAL READING 1. Guidelines for Obtaining Authorization of Compounds to be used in Meat and Poultry Plants, FS&I Service Agriculture Handbook #562, reprinted December 1990, U.S. Department of Agriculture, Washington, D.C. 2. Morawek, R. and Rhodes, R., Food Grade Lubricants and Their Applications, paper presented at S.T.L.E. Ann. Meeting, Dearborn, MI, April 17 to 20, 1978. 3. Code of Federal Regulations, Vol. 21, Parts 170 to 199, revised annually, U.S. Government Printing Office, Washington, D.C. 4. Food Engineering Magazine, Chilton Publication, Radmor, PA, (published 11 times per year). 5. Arbocus, G., Lubrication practices in food processing industries, CRC Handbook of Lubrication, Vol. 1, Booser, E. R., Ed., CRC Press, Boca FL, 1983, 359–371. 6. Brown, J., Power Transmission Design, Penton Publishing, Cleveland, OH, 39–42, Oct. 1991.

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31 Forest and Paper Products Frederick J. Villforth, III CONTENTS Key to Lubricants..............................................................................................................................309 Some Lubricant Recommendations in the Forest and Paper Products Industry....................311 References...........................................................................................................................................319

KEY TO LUBRICANTS HO — Antiwear Hydraulic Fluids Typically ISO VG-32, 46, 68, 100, 150, 220, antiwear hydraulic fluids made from high quality paraffinic base oils with excellent stability, containing additives to prevent rust, oxidation, and deposit formation, and to promote very rapid release of entrained air. Excellent wear protection is required for a variety of hydraulic pump metallurgy. CO — Circulating Oils for Centralized Lubrication Systems Typically ISO VG-32,46, 68, 100, 150, 220 lubricating oils made from high quality paraffinic base oils, with additives to prevent rust, oxidation and deposit formation, antifoamants. They must exhibit good water separation and filterability. RC — Residual Compounded Oils for Open Gears, Sprockets, Wire Rope, and Chains Heavy, adhesive-type lubricants typically meeting requirements of AGMA Lubricants 14R and 15R. These compounded residual oils can be semifluid to solid. They should provide excellent wear protection, prevent rust, resist water washout, and adhere to metal surfaces. GO-IEP — Industrial Gear Oils, Extreme Pressure Typically ISO VG 68, 150, 220, 460, and 680 oils which meet requirements for AGMA No. 2EP, 4EP, 5EP, 7EP, 8EP. A high quality industrial gear oil with effective extreme pressure additive, good oxidative and thermal stability, good antifoam and water shedding properties. GO-HY — Gear Oil for Hypoid Gears Typically SAE 80W-90 viscosity grade oils, demonstrating API GL-4/5 (shock loading) performance are used for hypoid gear differentials in the transportation industry. Oils should demonstrate excellent thermal stability, EP, antiwear and antifoam properties, demulsibility, rust and corrosion protection. GR-C — Coupling Grease Typically NLGI-0/1 greases, meeting requirements of AGMA CG-1, CG-2, or CG-3. Special grease for all types of couplings, formulated with special thickeners, high-viscosity base oil, and tackifier to prevent high speed separation and throw-off. Also provides extreme pressure and antiwear properties, superior resistance to water washout, corrosion, and rust protection. Suitable for use in continuous service up to 250°F and for short periods up to 325°F.

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GR-M — Multipurpose Industrial and Automotive, Extreme Pressure (EP) Grease Typically NLGI-2 greases, or NLGI-1 for some applications. Multipurpose grease, prepared from highly refined mineral oils and containing oil-soluble, nonlead extreme pressure additives. Provides rust and oxidation inhibitors, and resists water washout. Suitable for use in continuous service up to 250°F and for short periods up to 350°F. GR-S — Heavy Duty (HD) Industrial, EP, High-Temperature (HT) Operation Grease Typically NLGI-2 greases, or NLGI-1 for some applications. Grease with high dropping points, prepared from highly refined paraffinic base oils, with additives to provide superior rust and corrosion protection, resistance to water washout, oxidation stability, extreme pressure properties, and wear protection under high loads. Suitable for use in continuous service up to 325°F and for short periods up to 450°F. GR-P — HD Industrial, EP, High Load, High Water Exposure, HT Operation Grease Typically NLGI-1 grease with characteristics described for GR-S, but higher viscosity base oils (ISO 460) provide protection under high loads and under excessive water exposure conditions. Suitable for use from 0°F to continuous service up to 325°F, and for short periods up to 450°F. GR-NS — HD Industrial, EP, HT Operation, Nonstaining, Synthetic Base Oil Grease Typically NLGI-1 or 1½ grease, made with a synthetic base oil to provide a nonstaining grease. These types of greases have high dropping points (450°F+), provide the highest levels of EP and antiwear properties, superior rust and corrosion protection, resistance to water washout, and excellent oxidation stability. Suitable for use from -30°F to continuous service up to 350°F, and for short periods up to 450°F. GR-MG — Grease for Operating under Severe Water and Load-Carrying Conditions Typically an NLGI-2 grease made with high-viscosity base oil, with molybdenum and graphite added to provide high load-carrying capability and superior resistance to water washout. WR — Wire Rope Lubricant A wire rope compounded oil, usually semifluid and tacky, with a low pour point so that it remains pliable under most ambient conditions, adheres well to metal surfaces, provides wear and rust protection, and resists water washout. PT — Pneumatic Tool, and Other Air Cylinders, EP Lubricant Typically ISO VG 100 oils, blended specifically for air tools from stable base stocks. They contain extreme pressure and oiliness additives, have low fogging, good antifoam, rust and corrosion protection, low deposit-forming tendencies, and easy emulsification. PMO — Paper Machine Oil Paper machine oils are specially designed circulating oils for use in centralized lubrication systems of paper-making machines. These oils use highly refined base stocks with additives to provide excellent resistance to oxidation, rust and copper corrosion protection, resistance to deposit formation, excellent demulsibility, and antiwear and filterability properties. PMOs are typically available in ISO VG150, 220, 320, and 460. SCO — Steam Cylinder Oils and Worm Gear Lubricants Typically ISO VG-460, 680, and 1000, meeting requirements for AGMA 7 Comp, 8 Comp, and 8A Comp oils. These oils are compounded high VI oils that separate easily from condensate and have excellent lubricity, wetability, foam resistance, and rust protection. USDA/FDA — Personal Care Product Manufacture Lubricants Lubricating oils, greases, and hydraulic fluids that are approved by the U.S. Department of Agriculture as H-1, Lubricants with Incidental Contact, and meeting the U.S. Food and Drug Administration specification 21 CFR 178.3570. These lubricants must be the correct viscosity and provide the performance required by the machine in which they are used. API — American Petroleum Institute API classifications indicate a product quality level demonstrated by successfully passing performance tests in the type equipment in which the lubricant is to be used. Engine oil tests must be conducted according to the Chemical Manufacturers Association Product Approval Code of Practice. Oils must be licensed to use an API designation. Copyright © 1997 CRC Press, LLC.

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SAE — Society of Automotive Engineers SAE defines viscosity ranges of lubricants under various test conditions. These specifications do not define lubricant performance characteristics that may be required for the intended application. ISO — International Standards Organization Eighteen ISO Viscosity grades from 2 to 1500 are based on minimum/maximum kinematic viscosity (cSt) measured at 40°C. These grades do not define lubricant performance characteristics that may be required for the intended application. ILSAC — International Lubricant Standardization and Approvals Committee ILSAC issues specifications listing minimum performance standards for gasoline engine oils in a variety of engine and physical tests. Qualifying oils of 0W-, 5W-, and 10W-SAE viscosity grades can use the ILSAC “Starburst” Certification Mark. NLGI — National Lubricating Grease Institute A series of numbers classifying the consistency range of greases, in order of increasing hardness defined by an ASTM penetration test. This classification does not define grease performance characteristics that may be required for the application. AGMA — American Gear Manufacturers Association A series of lubricant specifications for industrial enclosed gear drives and industrial open gearing. Various applications include Rust and Oxidation Inhibited, Extreme Pressure (EP), Compounded Lubricants, and Residual Lubricants.

SOME LUBRICANT RECOMMENDATIONS IN THE FOREST AND PAPER PRODUCTS INDUSTRY

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(Note: After January 1996, API CD, CD-II, CE categories will no longer be licensed; some products may still be in the marketplace.)

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REFERENCES 1. LUBRICATION, Forest Products Industry-I, Logging and Lumber, Volume 80, No. 2, 1994; LUBRICATION, Forest Products Industry-II, Stock Preparation and Chemical Recovery, Volume 80, No. 3, 1994; LUBRICATION, Forest Products Industry-III, Paper Making and Finishing, Volume 80, No. 4, 1994 — Texaco Inc., New York. 2. Handbook of Paper Science: The Science and Technology of Papermaking, Paper Properties & Usage, 2 Volumes, 1980 and 1982, Ranee, H. E, Elsevier Scientific, New York. 3. Saltman, D., Paper Basics: Forestry, Manufacture, Selection, Purchasing, Mathematics and Metrics, Recycling, Kroeger Publishing, 1991, Melbourne, FL. 4. Timber Harvesting, 4th ed., American Pulpwood Association Staff, 1988, Interstate Printers and Publishers, Danville, IL.

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Machine Tools* Douglas W. Vallance

CONTENTS Introduction..................................................................................................................................................320 Machine Tool Construction......................................................................................................................320 Environmental Considerations.................................................................................................................322 Particulate Contamination in Hydraulic and Lubrication Systems....................................................322 Machine Condition Monitoring................................................................................................................325 Typical Lubricant Delivery Systems.........................................................................................................326 Lubricants Classification............................................................................................................................327 References.....................................................................................................................................................331

INTRODUCTION Machine tools are relied upon heavily by manufacturing industries for production of parts and ultimately finished goods. Machine tools include a diverse array of types, such as milling machines, lathes, grinders, planers, broaches, drills, and a wide variety of multipurpose and specialized equipment. Computer numerical control and automatic work handling are common. The quality of a machine tool is inevitably judged by its speed and precision. Lubrication facilitates this performance through the reduction of friction and its two detrimental attendant effects, wear and thermal instability. For simplification, ease of maintenance, and increased reliability, there has been a concerted attempt to reduce the number of different lubricants required in a given machine. Since a great multiplicity of components all perform different functions on even the simplest machine, however, it is not uncommon to require several different products.

MACHINE TOOL CONSTRUCTION A brief discussion of the types of materials used in machine tools may promote a better understanding of their lubrication requirements. STRUCTURAL MEMBERS The degree of rigidity required to minimize deflection dictates a somewhat massive structure and, consequently, stress levels are insignificant in beds, bases, slides, housings, and other structural members. Cast iron serves in most applications. Fortunately, gray cast iron possesses excellent vibration dampening capacity, and the free graphite particles provide some inherent lubrication.

* This section is revised and updated from J. E. Denton, Machine Tools, in CRC Handbook of Lubrication, Vol. 1, CRC Press, 1983, 167-180.

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Fabricated mild steel weldments are frequently used in very large machines where a casting is not practical. While fabricated steel weldment design has progressed to the point where dynamic vibration dampening is equivalent to cast iron, bearing properties of mild steel are unsuitable, and attached bearing strips of another material are required. Aluminum castings take advantage of their low density and reduced inertia in high-speed equipment-supporting members such as brackets, housings, and arms. Since bearing properties of aluminum are marginal, bushing inserts of bronze or some other material are sometimes required. SLIDEWAYS Where cast iron is used as the base or structural material, the slideways are usually made integral with the casting. When steel fabrications are used, slideways of hardened steel or specially alloyed and heat-treated cast iron strips may be attached to the bed. By far, however, the major application in plain slideways (i.e., flat surface vs. flat surface) involves the use of pearlitic gray cast iron. In numerically controlled equipment, two different systems are used to achieve the lowest possible friction: (1) plastic bearing inserts based on polytetrafluroethylene (PTFE) and (2) antifriction linear roller bearings on hardened steel ways. POWER TRANSMISSION COMPONENTS Spindles, shafts, gears, and other power transmission components are generally made of heat-treated steel. Hardened and tempered medium carbon alloy steel such as AISI 4150 is common. Carburized and hardened steels such as AISI 8620 are frequently used where high-bearing contact stress is encountered. Cams, guides, rails, and other members subject to extreme wear are sometimes made from highly alloyed tool steels such as AISI A-2, M-2, or similar wear-resistant grades. BEARINGS Antifriction ball or roller bearings are used extensively in machine tools. Through-hardening and case-hardening steels are commonly used in these rolling element bearings. Hydrostatic bearings, also frequently employed in machine tools, use cast iron, tin bronzes, oil impregnated sintered bronzes, tin babbitt, and occasionally nylon. SURFACE COATINGS Typical coatings used on machine tool components to enhance friction characteristics, to increase wear resistance, or to improve corrosion resistance include anodizing, hard chromium, phosphating, and nickel-based coatings with and without fillers. Under low-unit load conditions, anodized coatings perform well on aluminum to prevent galling and to increase wear and corrosion resistance. Hard chromium deposits are sometimes specified for enhancement of frictional characteristics or resistance to fretting. They are frequently used on shafts that run against elastomeric lip seals. Hard chromium shows a remarkable tendency towards nonwetting and does not maintain a continuous fluid film of lubricant. Phosphate coatings are used primarily as attrition coatings to provide resistance to pick-up and scoring during the wear-in phase on new equipment. The crystalline coating is sacrificially worn away, with the wear debris serving to prevent metal-to-metal contact. Phosphate coatings are typically absorbent and provide a good base for retention of lubricants. Manganese phosphate coatings are the preferred type for this application. Electroless nickel systems provide hard, corrosion-resistant surfaces of uniform thickness that resist wear and abusive environmental conditions. Very low surface friction can be achieved through addition of second phase particles such as PTFE-based compounds.

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ENVIRONMENTAL CONSIDERATIONS TEMPERATURE FLUTUATIONS Temperature fluctuations can have a drastic effect on the precision of a machine tool. Thermal gradients existing within a structure can result in very complex geometric distortions. Dissimilar metals such as steel, cast iron, bronze, aluminum, and stainless steel can introduce further complexities owing to their different coefficients of expansion. Temperature changes and subsequent thermal expansions may significantly affect precision-bearing preloads as well. Ideally, machine tools operate in benevolent surroundings for the comfort of the operator. Where very high-precision work is being done, it may be necessary to provide a controlled constant temperature environment. Most machine tool lubricants operate in a “room” temperature ambient at or near 40°C. Thermal gradients may arise from operational effects as well as changes in ambient conditions. Heat is generated within the machine itself by such sources as friction, electrical motors, fluid shear, and gas compression. Normal operating procedure involves a warm-up period to allow the machine to reach some degree of thermal equilibrium before beginning actual production. The hydraulic system represents the dominate source of internally induced heating. Fluid shear contributes to most of the bulk system temperature rise. Another real heat source is adiabatic compression of entrained air bubbles to intense elevated temperatures as the fluid passes through the hydraulic pump.1 High-production equipment, where heat generation exceeds heat loss, requires supplemental cooling by chilled water or refrigerated heat exchangers. This is necessary both to control thermal distortion of the equipment and to protect the hydraulic fluid from thermally induced degradation. These devices are normally set to maintain a bulk oil temperature of 40°C (104°F). MOISTURE CONSIDERATIONS Moisture, both as condensation from humid air and as contamination from water-based cutting fluids, can have a very deleterious effect on the performance and life of the lubricant. If the housing is not properly vented, cooling-off periods when the machine is shut down can generate substantial negative pressure inside the machine to draw cutting fluid past otherwise tight seals. Normally, water is undesirable as a dispersant in machine tool lubricants. The maximum oil temperature reached in service is not sufficiently high to drive off dispersed water, and sludging or rusting problems may result. It is preferable that entrained water in the system be allowed to separate out in a location where it can be drained off. The oil should contain a rust inhibitor to prevent rusting of component parts. Water can also have a detrimental effect on the life of rolling element way bearings and internal ball and roller bearings when the lubricant in use inadvertently becomes cross contaminated with water-based metal working fluids commonly employed during machine tool usage. “A concentration of water of 0.01% is enough to decrease rolling element bearing life to half its original value.”2,3

PARTICIPATE CONTAMINATION IN HYDRAULIC AND LUBRICATION SYSTEMS Solid particulate matter can accumulate in lubricants, causing excessive wear, seizing of hydraulic components, and degradation of the lubricant. The typical radial clearance between the spool and bushing in a hydraulic servo valve is 10 µm (0.0004 in. or 10 microns). This means a 10-µm particle can seize the valve, and 5- to 10-µm particles can promote severe wear, cause sticking and subsequent reduced performance and reliability. The lower limit of resolution of the unaided human eye is approximately 40 µm (0.0016 in.). Obviously, it is impossible to visually determine when a fluid is clean for numerically controlled servo system requirements without the aid of a microscope.

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Hard, solid particles of size greater than the elasto-hydrodynamic film thickness in a rolling element can decrease bearing fatigue life 80 to 90%,3,5 while extremely clean lubricant can extend bearing life 15 to 30 times past normally calculated life.3,6 SOURCES OF CONTAMINATION Most particulate matter that finds its way into hydraulic systems comes from three basic sources: residual debris from manufacturing operations, contamination from external sources, and internally generated wear debris. Typical examples of built-in contamination include machining chips, core sand from castings, rubber fragments from hoses and seals, lint from shop towels, rust, and weld scale. Contaminates from external sources include airborne grinding grit, paint spray, dust, and lint, which enter the system through lack of adequate filtration in air breathers, open fill caps, and through seals on actuators. Even though oil as refined may be very low in particulate level, oil from drums is invariably contaminated unless special filtration procedures are followed during machine reservoir additions. Internally generated contamination may consist of wear debris, paper fibers from deteriorating filter elements, sheared “O” rings, and oil degradation products. FILTRATION The practice of using the hydraulic fluid as a general purpose lubricant for gear trains, clutches, and bearings requires a properly designed filtration system. During break-in periods of operation, excessive amounts of wear debris may be generated which can result in self-destructive operation. During assembly of a new machine, it is common practice to substitute a temporary manifold block for critical servo valves and flush the system by recirculating the hydraulic fluid with appropriate external filtration. When it has been determined by periodic particle counts that the system is clean, the temporary flushing manifolds are removed and the servo valves installed. This procedure prevents clogging and seizing of these critical valves during the period when the built-in contamination level is highest. In general operation, machine tools require continuous filtration to remove airborne contaminants and internally generated wear debris. Two general categories of filtration use either absolute or nominal type filters. Absolute filters generally require a beta ratio of 75 or better, where the beta ratio is given by:7

Filter efficiency is then calculated as follows:7

These efficiencies may be calculated during multipass testing. Filters rated to a nominal micron value will usually allow passage of a portion of particles greater in size than the filter rating, but these filters usually increase in efficiency as they become dirtier, which allows for collection of smaller size particles. Absolute filters at pump outlet lines are used to protect critical systems from debris in the event of a pump failure. Filtration can also be used in return lines or in bypass systems such as those used on oil coolers. Return line and bypass system filters can be generally rated to a much lower operating pressure, thereby reducing cost. Copyright © 1997 CRC Press, LLC.

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A pressure differential sensor gives a signal when a predetermined level of filter loading is reached, and the filter automatically shifts into by-pass. It is essential to change the filter promptly when this occurs to protect system components. In very critical systems the filter load signal may actually shut down the equipment. DETERMINATION OF CONTAMINANT LEVEL Several techniques have been devised to determine the contamination level in hydraulic systems. Perhaps the most direct method is to filter a known volume of the fluid through a membrane filter that retains all of the particles on its surface. These particles can then be examined microscopically to identify their origin and counted to yield a quantitative contamination level. The disadvantage of this technique is that it is very tedious and time-consuming. A semiquantitative method sometimes used is to filter a known volume of sample through a membrane-type filter and compare the resulting appearance to a series of standard filter samples. This method can yield misleading information, since evaluation depends upon the intensity of color rather than particle size or distribution. Direct reading ferrography subjects a stream of fluid to a magnetic field gradient and measures decreased light penetration at known positions of the field. This decrease in light passing through the stream is correlated to the amount of ferrous particles in the oil. Analytical ferrography also subjects a thin stream of fluid to a magnetic field gradient. Particles are precipitated from the fluid as a function of size and magnetic permeability to produce a sample “streak.” These samples may be fixed for retention as a permanent record. They may also be evaluated qualitatively by microscopic examination. Probably the most appropriate method for rapid particulate level determination is automatic counting by light beam interruption. In this technique, the sample fluid is made to flow through a narrow aperture past a window. A laser light beam passes through the window and fluid and falls upon a photo tube. When an opaque particle carried by the fluid stream passes through the light beam, an electric signal is generated by the photo tube, which is proportional to the size of the particle. By suitable circuitry these signals may be classified and counted to yield a particle size vs. count distribution curve. Entrained air must be removed from the sample prior to particle counting as gas bubbles will also be counted as particles. Vacuum degassing is one common way to remove entrained air from oil samples, or slight pressurization will cause the air bubbles to dissolve in the oil. Automated techniques to measure oil flow through calibrated membranes are available which have shown good correlation with light- or laser-type particle counters. Capacitative methods are also used with commercial success to determine contamination levels in hydraulic and lubrication samples. Sampling techniques for oil analysis are extremely important. Sampling locations and conditions should remain constant over time. Clean sample bottles and draw tubes should be utilized. Sampling should occur away from the sides or bottom of reservoirs. Sampling ports in hydraulic lines should be positioned in areas of turbulent flow. Samples should be agitated or shaken prior to measuring particle contamination. Failure to take these precautions may result in inaccurate results and improper corrective action. Standards in current use with hydraulic fluids for expression of particle size and distribution tests include A1A NAS 1638 and ISO 4406 as shown in Table 1. The current industry standard in machine tool oil cleanliness is to report a two-digit ISO cleanliness code based on the number of particles greater than 5 and 15 as an ISO range5µm/range15µm. For example, an oil sample containing 2000 particles greater than 5µm /ml fluid and 30 particles greater than 15 µm/ml fluid would be reported as an ISO 18/12 cleanliness code. This code can also be reported with particles greater than 2 µm for increased resolution, i.e., an ISO 24/18/12 code. The methodology used should be reported along with the result code. Oil cleanliness levels to be maintained should be determined on an individual basis due to the wide variety of systems in use. Hydraulic servo control systems require the greatest cleanliness,

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while simple hydraulic power transmission systems may be more tolerant to particle contamination. Hydraulic pump and valve manufacturers recommend oil cleanliness values in their literature or will provide values upon request. Ten micron-nominal rated filters have been found to maintain cleanliness levels around 18/15 (5- and 15-µm size, respectively). This level is generally more contaminated than recommended for machine tool hydraulic pumps and valves. Use of 3- to 5-µm filters can produce levels 12/8 and below which is a fairly clean range for machine tool operation. These levels can vary dramatically, depending on the effectiveness in prevention of ingress of external contamination. Table 2 reports some general guidelines for machine tool hydraulic cleanliness levels. Increased cleanliness has been shown to be effective in reducing internal machine tool component wear and increasing long-term reliability.

MACHINE CONDITION MONITORING Machine condition monitoring is enjoying more widespread use as it has become more exact and cost-effective in helping to prevent catastrophic failure, reduce down time, decrease maintenance costs, and increase reliability. Machine condition monitoring is in part accomplished through analysis of wear particles carried in lubricants during normal operating service. Oil or grease samples are taken periodically and characterized for amounts of wear particles along with morphologies and potential sources to trend the machines’ operating characteristics. When abnormal Copyright © 1997 CRC Press, LLC.

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particles are present or seen to be significantly increasing, corrective actions are recommended. These can usually reduce downtime for machine repair and often eliminate secondary failures due to system contamination. Instrumentation for in-house analysis and commercial laboratories which provide machine condition monitoring services are becoming more common. At the same time the machine condition is checked, evaluation of the lubricant for properties which make it effective (viscosity, total acid number, etc.) extend the lubricant use as long as possible.

TYPICAL LUBRICANT DELIVERY SYSTEMS SYSTEM DESIGN A detailed analysis of recommended practice when designing a lubrication system for a machine tool is given in ASLE Standard Number 70–4.8 The machine tool user must be furnished with a lubrication schedule including a schematic drawing showing fill points, a description of the lubricant to be used, the capacity of the system, and the lubrication intervals to be observed. ISO 5169 — Machine Tools — Presentation of Lubrication Instructions provides an effective way to present this information as illustrated for a vertical machining center in Figure 1. Plates containing this information are attached to individual machines. LUBRICATION SYSTEM ELEMENTS There is a trend in the machine tool industry to simplify lubrication requirements through fewer types of lubricants, lubed-for-life units, and automatic lube delivery systems. ISO 5170 — Machine Tools — Lubrication Systems provides an international standard for system classification and specifications regarding components of these systems. Hand oilers are used only in the simplest equipment because lubrication can only be applied to accessible points, and the operator must be relied upon to attend conscientiously to this chore. Sight feed oilers represent some improvement over hand oilers in that a reservoir is provided along with a metered rate of application. Gear boxes and transmissions usually involve partial immersion or splash lubrication to lubricate gears and bearings. In cascade or flood systems, a continuous stream of lubricant is pumped over enclosed components such as spindles and gear trains mounted in housings. Frequently, supply oil from the hydraulic system is used for this purpose. Centralized forced feed systems incorporated in virtually all high-production machines consist of a reservoir, a pump, and delivery lines to adjustable metering valves at the various points requiring lubrication. This type of system can also utilize the hydraulic oil supply. The advantage of a central system is that a constant supply of lubricant is delivered even to points inaccessible to the operator whenever the machine tool is in operation, without the need for attention by the operator.

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FIGURE 1 Schematic illustrating effective lubrication instructions and schedule for machine tool lubrication maintenance. (From Cincinnati Milacron, Cincinnati, OH. Saber Vertical Machining Center lubrication plate. With permission.) Mist lubrication is a specialized method which pneumatically atomizes the lubricant into an aerosol for delivery, usually to ball bearings. The aerosol particles coalesce into larger droplets on the bearing surfaces to provide the desired lubrication. The air stream that carries the droplets also serves to cool the bearing. When operating at moderate to high velocity, a ball bearing can act as a pump to force a liquid stream back out of the bearing so that flow of the lubricant through the bearing is impeded. Mist lubrication helps to overcome this tendency. Air/oil systems are becoming more frequently utilized as the trend for increased spindle speeds continues. Using a controlled oil injection system to deliver precise amounts of oil to individual bearings at controlled intervals, these systems allow for better delivery control than does mist lubrication.

LUBRICANTS CLASSIFICATION Lubricants are classified by the ISO 3498 system shown in Table 3, which consists of letters identifying the lube function followed by a number representing the viscosity in centistokes at 40°C. For example, CKC 320 is a gear lube having a viscosity of 320 cSt with extreme pressure and antiwear properties. The following list describes some of the functions and special requirements of the lubricants used in machine tools. Copyright © 1997 CRC Press, LLC.

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GEAR OILS These consist of mineral oils with extreme pressure (EP) additives designed to react with the metal surfaces to enhance protection of highly loaded gears. Extreme pressure additives vary in the degree of reactivity from mild to very aggressive. Some EP additives are corrosive to copper alloys such as those used in worm gear sets, and caution should be taken in the selection of such products. Many lightly loaded high-speed gears are lubricated with hydraulic or spindle lube. Typical viscosity ranges for gear lubes are 68, 150, 320, and 460 cSt at 40°C. Worm gears generally utilize heavier viscosities and less aggressive extreme pressure additive packages. Worm gears primarily have a sliding interface which requires the oil to exhibit good lubricity in order to help reduce friction. While in the past compounded oils were utilized, the current trend is for use of synthetic lubricants in worm gear applications. SPINDLE OILS These are highly refined mineral oils with good oxidation resistance. Viscosities are relatively low, since the parts lubricated are usually operating at high velocities. Mechanisms involved in spindle lubrication include antifriction and hydrostatic bearings, electromagnetic clutches, and lightly loaded gears. Typical viscosity ranges for spindle oils are 2, 10, 22, and 32 cSt at 40°C. WAY LUBRICANTS These require special antistick-slip and tackiness additives for the lubrication of plain bearing slideways. When slideways operate at low relative motion, in the order of 1 to 5 cm/min (0.5 to 2.0 in./min), two coefficients of friction are significant. The static or break-away coefficient of friction represents the force required to put the system into motion. The kinetic coefficient of friction represents the force required to maintain motion. When the static coefficient of friction is sufficiently greater than the kinetic coefficient of friction, motion may progress in a series of jumps called stickslip.9 The additive should ensure that the ratio of the static to kinetic coefficient of friction is less than unity and preferably as low as 0.80. The function of the tackiness additive is to impede runoff of the lubricant, especially on vertical surfaces. Way lubricants are normally supplied in viscosities of 32, 68, and 220 cSt at 40°C. HYDRAULIC OILS These products consist of two basic families: conventional rust- and oxidation-inhibited oils and antiwear hydraulic oils. While their basic function is to transmit fluid power, they are also called upon to lubricate pumps, valves, and other system components. Combination systems are common whereby the hydraulic oil also serves as a spindle oil and way lubricant. Conventional rust- and oxidation-inhibited hydraulic oils have proven to be very stable and are useful at system pressures up to 90% of the rated capacity of the pumps. In high-pressure systems, especially those employing vane and rotor type pumps, antiwear additives may be required. The original zinc dithiophosphate antiwear additives were particularly unstable, and moisture, copperbearing metals, and elevated operating temperatures catalyzed oxidation, sludging, and general degradation. Nonmetal sulfur-phosphorus type antiwear additives and very stable zinc based additives have now been developed that provide the required resistance to oxidation and sludging while imparting very good antiwear properties. Only the best quality antiwear lubricants should be used, since in machine tools (a) copper metals are used, (b) moisture will probably be present, and (c) high-localized temperatures are generated. The preferred viscosity ranges for hydraulic fluids are 32, 46, 68, and 150 cSt at 40°C. The reasonably discriminating Standard Test Method for Thermal Stability of Hydraulic Oils (ASTM D 2070) consists of heating 200 ml of oil containing polished copper and steel rods to a temperature of 135°C (275°F) for 168 hr. Subsequent filtration of the precipitated sludge and

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evaluation of the corrosion of the copper and steel rods form the basis of determination of oil stability. Products which form less than 25 mg of sludge per 100 ml of oil are considered satisfactory, between 25 and 100 marginal, and over 100 mg unsatisfactory. GREASES As lubricated-for-life components become more prevalent in the machine tool industry, greases are seeing more widespread use. Greases used in permanently lubricated applications need to exhibit long life properties and good durability. Increasingly higher speed spindle assemblies necessitate the use of greases employing lower viscosity base oils. Multipurpose greases are used in machine tools to lubricate plain and rolling bearings and miscellaneous parts where runoff of a once-through system cannot be tolerated. While many metallic soap types are available, the greatest volume involves lithium-base greases, which generally have excellent resistance to softening upon working and good resistance to water washing. It should be emphasized that different types of greases are not necessarily compatible. Soda-based greases, for example, should not be mixed with lithium-based greases or thinning will result. Polyglycol-based lubricants should not be exposed to mineral oils. Extreme pressure greases containing EP additives are very useful in lubrication of shock-loaded ball and roller bearings. Greases containing molybdenum disulfide are also useful for extreme pressure sliding applications, especially where very low speeds are prevalent. In general, greases used in machine tools should not contain fillers such as clay, mica, or asbestos. Grades ranging from NLGI 00 (semifluid grease) to NLGI 3 (high consistency) are seeing more widespread use. For general purpose applications, base oil viscosity generally ranges from ISO 100 to 220 grades and consistencies of NLGI 1 or 2 are utilized. Worm gears may be lubricated with synthetic, semifluid greases (NLGI 00) with base oil viscosities from ISO 220 to 680, depending on loading and speed. These worm gear lubricants must exhibit good lubricity to combat excess heating due to high friction and sliding. Lithium and barium complex thickened synthetic greases with low base oil viscosity (near ISO 32) see wide use in high-speed spindle bearing applications. These usually utilize an NLGI 2 or 3 consistency to help keep the small amount of lubricant required in place during service.

REFERENCES 1. Roberton, R. S. and Allen, J. M., A study of oil performance in numerically controlled hydraulic systems, paper presented at the National Conference on Fluid Power, Mobil Oil Corp., New York, 1974. 2. Schatzberg, P. and Felsen. I. M., Influence of water on fatigue failure location and surface alteration during rolling contact lubrication, J. Lubr. Technol., ASME Trans., F, 91, 1 (1969), 301–307. 3. Ioannides, E. and Jacobson, B., Dirty lubricants — reduced bearing life, Ball Bearing J. Special ′89, 22–27. 4. Okamoto, J., Fujita, K., and Yoshioka, T., Effects of solid particles in oil on the life of ball bearings, J. Mech. Eng. Lab., Tokyo, 26(5), Sept. 1972. 5. Fitzsimmons, B. and Cave, B. J., Lubricant contaminants and their effects on bearing performance, Society of Automotive Engineers, 750583 (1975). 6. Tallian, T, Prediction of rolling contact fatigue life in contaminated lubricant, II. Experimental, ASME Trans., July 1976, J. Lubr. Technol., pp. 384–392. 7. Fitch, J., Oil Analysis and Proactive Maintenance, Proactive Maintenance Institute, 1995-1 ed. 8. Lubrication Standard for Mass Production Industry, Machine Tools and Related Equipment, Standard No. 70-4, American Society of Lubrication Engineers, Park Ridge, IL, 1973. 9. Rabinowicz, E., Friction and Wear of Materials, John Wiley & Sons, New York, 1966.

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Mining Excavation Equipment Lubricant Guidelines Charles D. Barrett

CONTENTS Introduction....................................................................................................................................332 Gear Oil — Single Viscosity Grade; Extreme Pressure (EP)................................................333 Gear Oil — Multiviscosity Grade; Extreme Pressure (EP)....................................................334 Synthetic Gear Oil.........................................................................................................................335 Circulating/Hydraulic Oil: Antiwear...........................................................................................336 High VI Circulating/Hydraulic Oil: Antiwear..........................................................................337 Multipurpose Grease; Extreme Pressure (EP)..........................................................................338 Open Gear Lubricant....................................................................................................................339 Wire Rope Lubricant.....................................................................................................................340

INTRODUCTION These lubricant guidelines have been developed to assist users of mining excavation machinery to select lubricants for shovels, draglines, blast hole drills, front end loaders, haul trucks, etc. When used in accordance with their suppliers’ recommended practices, these guidelines should provide a minimum level of satisfactory equipment performance. The use of premium lubricants that exceed these guidelines and performance requirements is suggested. Despite their higher initial purchase cost, these higher performance products routinely generate a significant return on investment over the useful service life of mining excavation machinery. The minimal lubricant performance characteristics that follow do not attempt to address the multitude of individual federal, state, and local safety, health, environmental, and disposal issues that may affect their use in mining excavation machinery applications. It is the responsibility of the machine user to verify that lubricants in service meet all applicable safety, health, and environmental regulations. The eight most common lubricant types used in mining excavation machinery are covered by these guidelines. Other products such as cam and slide and multiservice lubricants are not included. These products are either application specific for individual manufacturers, or their use is not universally accepted by the majority of equipment manufacturers. The manufacturer of the mining excavation machinery should be consulted prior to selecting a suitable lubricant for applications requiring these types of products.

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TABLE 1 GEAR OIL — SINGLE VISCOSITY GRADE; EXTREME PRESSURE (EP) SCOPE This specification covers premium single-grade gear oils produced from refined mineral oilbase stocks, and compounded with extreme pressure additives for high load-carrying ability, used in sump (splash), pressure circulating, or total loss applications. These products are intended to lubricate gears, bushings, plain and rolling element bearings, and other components enclosed in oil-tight housings, and operating under low speed, high-torque, high-shock load, and/or high-speed, low-torque conditions. PHYSICAL CHARACTERISTICS These gear oils are to be manufactured from highly refined mineral oil-base stocks. These lubricants must maintain their viscosity and performance characteristics throughout their intended service life. They may not contain viscosity or viscosity index enhancers that are subject to breakdown under extended shearing conditions. EP or other friction modifiers used in the formulation of these lubricants must be fully oil soluble, and/or be in a stable colloidal suspension. MINIMUM PERFORMANCE REQUIREMENTS

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TABLE 2 GEAR OIL — MULTIVISCOSITY GRADE; EXTREME PRESSURE (EP) SCOPE This specification covers premium multiviscosity gear oils, produced from refined mineral, synthetics, or blends of synthetic and mineral-base fluids. Materials furnished under this specification are intended to lubricate gears, bushings, plain and rolling element bearings, and other components enclosed in oil-tight housings, and operating under low-speed, high-torque, highshock load, and/or high-speed, low-torque conditions. They may be used in sump (splash), pressure circulating, or total loss applications. Materials furnished under this specification are intended for use when extremes in ambient temperatures may dictate seasonal changes to different viscosity grades. PHYSICAL CHARACTERISTICS These gear oils shall be a homogeneous petroleum product, a synthetically prepared product, or combination of the two types of products. These lubricants may not contain viscosity or viscosity index enhancers that are subject to breakdown under extended shearing conditions. These lubricants must have chemical and physical stability to maintain viscosity and performance characteristics throughout their intended service life. Any EP or other friction modifiers used in the formulation of these lubricants must be fully oil soluble, and/or be in a stable colloidal suspension. MINIMUM PERFORMANCE REQUIREMENTS

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TABLE 3 SYNTHETIC GEAR OIL SCOPE This specification covers premium gear oils, produced from synthetic or semisynthetic base stocks, intended to lubricate enclosed gears, bushings, plain and rolling element bearings, and other components enclosed in oil-tight housings and operating under low-speed, high-torque, high-shock load. They may be used in sump (splash), pressure circulating, or total-loss applications. Materials furnished under this specification are primarily intended for use in applications where bulk oil temperatures routinely range from 77°C/170°F to 121°C/250°F. They may also be beneficial in applications subject to arctic conditions. While materials furnished under this specification provide extreme pressure protection, there is no requirement that they contain EP additives to achieve this property. PHYSICAL CHARACTERISTICS These gear oils must be formulated with full synthetic or semisynthetic base fluids. These lubricants must have chemical and physical stability to maintain viscosity and performance characteristics throughout their intended service life. They may not contain viscosity or viscosity index enhancers that are subject to breakdown under extended shearing conditions. EP or other friction modifiers used in the formulation of these lubricants must be fully oil soluble, and/or be in a stable colloidal suspension. MINIMUM PERFORMANCE REQUIREMENTS

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TABLE 4 CIRCULATING/HYDRAULIC OIL: ANTIWEAR SCOPE 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. PHYSICAL CHARACTERISTICS These lubricating oils must be manufactured from highly refined mineral oil base stocks. They must have chemical and physical stability to maintain viscosity and performance characteristics throughout their intended service life. These lubricants may not contain viscosity or viscosity index enhancers that are subject to breakdown under extended shearing conditions. All additives used in the formulation of these lubricants must be fully oil soluble. MINIMUM PERFORMANCE REQUIREMENTS

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TABLE 5 HIGH VI CIRCULATING/HYDRAULIC OIL: ANTIWEAR SCOPE This specification covers premium high VI (viscosity index) circulating oils produced from highly refined mineral oils, full synthetic base stocks, or semisynthetic blends. They are compounded with antiwear additives for high load carrying ability. Materials furnished under this specification are primarily intended for use in hydraulic systems operating within an ambient temperature range of -40°C/-40°F to 66°C/150°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. PHYSICAL CHARACTERISTICS These lubricating oils must be manufactured from highly refined mineral oils, full synthetic base stocks, or semisynthetic blends. They must have chemical and physical stability to maintain viscosity and performance characteristics throughout their intended service life. These lubricants may not contain viscosity or VI enhancers that are subject to breakdown under extended shearing conditions. All additives used in the formulation of these lubricants must be fully oil soluble. MINIMUM PERFORMANCE REQUIREMENTS

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TABLE 6 MULTIPURPOSE GREASE; EXTREME PRESSURE (EP) SCOPE This specification covers premium multipurpose, extreme pressure greases with essential properties and characteristics that make them suitable for use in diversified applications through a wide range of ambient temperatures. Both mineral and synthetic base oils may be used to formulate greases that comply with this specification. Materials furnished under this specification are intended to lubricate bushings, plain and rolling element bearings, and miscellaneous items. They may be applied by hand, hand pressure guns, pneumatic or electric pressure guns, or centralized lubrication systems of the single line parallel, single line progressive series, and/or dual line progressive construction. PHYSICAL CHARACTERISTICS These lubricating greases shall be manufactured with a lithium 12 hydroxystearate, lithium complex, or other thickeners, provided care is taken by the end user to avoid grease incompatibility with other lubricants or equipment components. They must contain extreme pressure additives, and be formulated to resist oxidation, corrosion, separation, and water washout. These greases must have chemical and mechanical stability to maintain consistency and performance characteristics throughout their intended service life. The NLGI grade of the grease must be appropriate for prevailing ambient temperature range. Class 3 grease is the preferred grade for all temperature ranges. However, at lower ambient temperature ranges, or where the application dictates, class 1 or 2 products may be recommended by the lubrication supplier. MINIMUM PERFORMANCE REQUIREMENTS

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SPECIAL CONSIDERATIONS The material furnished under this specification must be dispensable through the distribution lines of a centralized lubrication system to the most remote application point, at the lowest anticipated operating temperature. It must not plate or plug components of the centralized lubrication system, such as injectors, metering blocks, or spray nozzles. When the material furnished under this specification is formulated with solid film additives, particle size must not exceed 10 µm.

TABLE 7 OPEN GEAR LUBRICANT SCOPE This specification covers open gear lubricants with highly fortified blends of viscous fluids combined with additives to form stable, long-lasting, high load-carrying, wear-resistant films that lubricate under boundary conditions. Materials furnished under this specification are intended to lubricate open gears, racks, bushings, rails, rollers, dipper handles, and propel mechanisms. They are intended to be dispensed intermittently by centralized lubrication systems of the single line parallel, single line progressive series, and/or dual line progressive construction. This specification covers all grades of open gear lubricants that are usable, from -46°C/-50°F to 49°C/120°F. The particular grade and product selected must perform within the specific temperature range in which it is utilized and will depend on climate, application, performance, and regulatory requirements. Materials may be semifluid asphaltic compounds or semifluid pastes. PHYSICAL CHARACTERISTICS These open gear lubricants must have excellent adhesive and cohesive qualities, must not chip or throw off, provide sufficient film thickness to prevent metal-to-metal contact between applications under all operating conditions, must have excellent water-resistant and rust-preventive qualities, and have retarded dripping qualities for operation over wide temperature ranges. These open gear lubricants must be specifically formulated to protect surfaces, reduce wear, and provide normal component service life under all anticipated operating conditions. MINIMUM PERFORMANCE REQUIREMENTS

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SPECIAL CONSIDERATIONS The material furnished under this specification must be dispensable through the distribution lines of a centralized lubrication system to the most remote application point, at the lowest anticipated operating temperature. It must not plate or plug components of the centralized lubrication system such as injectors, metering blocks, or spray nozzles. Should the material furnished under this specification contain a diluent to improve dispensability, special care must be exercised to ensure its compatibility with all centralized lubrication system components, i.e., gaskets, O-rings, vent valves, etc. Diluent-containing lubricants furnished under this specification must conform to the Performance Requirements after the diluent has evaporated. When the material is formulated with solid film additives, particle size must not exceed 10 µm.

TABLE 8 WIRE ROPE LUBRICANT SCOPE This specification covers wire rope lubricants manufactured from mineral oil or synthetic base stocks, or a blend of mineral oil and synthetic base stocks. These materials are intended to lubricate large-diameter wire ropes such as those used for hoist and/or drag functions on draglines or shovels. Materials furnished under this specification are intended to be ideally dispensed by centralized lubrication systems of the single line parallel, single line progressive series, and/or dual line progressive construction. However, they are also suitable for application by pressure chamber, brush, or dip methods. This specification covers all grades of wire rope lubricants that are usable from -46°C/50°F to 49°C/120°F. The particular grade and product selected must perform within the specific temperature range in which it is utilized and will depend on climate, application, performance, and regulatory requirements Materials furnished under this specification may contain a diluent to enhance penetration to the rope core, or improve sprayability. PHYSICAL CHARACTERISTICS These wire rope lubricants must have excellent adhesive and cohesive qualities, must not chip or throw off, have excellent water resistant and rust-preventive qualities, and be 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/or solid film additives to improve film strength and control fretting and rubbing friction during operation.

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MINIMUM PERFORMANCE REQUIREMENTS

SPECIAL CONSIDERATIONS The material furnished under this specification must be dispensable through the distribution lines of a centralized lubrication system to the most remote application point, at the lowest anticipated operating temperature. It must not plate or plug components of the centralized lubrication system such as injectors, metering blocks, or spray nozzles. When the material is formulated with solid film additives, particle size must not exceed 10 µm. Should the material furnished under this specification contain a diluent to improve dispensability, special care must be exercised to ensure its compatibility with all centralized lubrication system components, i.e., gaskets, O-rings, vent valves, etc. Diluent-containing lubricants furnished under this specification must conform to the Performance Requirements after the diluent has evaporated.

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Refrigeration and Air Conditioning Lubricants Kenneth C. Lilje, Thomas E. Rajewski, and Edan E. Burton

CONTENTS Compressor Lubrication...................................................................................................................................342 Lubricant Requirements....................................................................................................................................343 Lubricant Types...................................................................................................................................................347 Retrofit of Existing Systems............................................................................................................................353 References.............................................................................................................................................................354

Lubricants play a key role in the success of refrigeration applications. The functions of the lubricant include reducing friction and wear, removing the heat of compression and providing sealing of the compressor. The increasing complexity of compressor system requirements, as well as the shift to non-ozone-depleting refrigerants, is affecting a change from traditional mineral oil refrigeration lubricants to synthetic lubricants. The wide range of properties available with synthetic lubricants allows for the customization of fluids to meet the specific needs of various refrigeration systems.

COMPRESSOR LUBRICATION LUBRICATION MODES Compressors generally operate under hydrodynamic and/or elastohydrodynamic-hydrodynamic lubrication conditions. Boundary lubrication is occasionally experienced under severe operating conditions such as start-up and high pressure operation.1 Lubrication under hydrodynamic or elastohydrodynamic conditions can usually be achieved using a correctly chosen lubricant alone. Antiwear additives are often required to protect metal surfaces exposed to prolonged boundary lubrication conditions. COMPRESSOR TYPES Compressor designs for refrigeration applications include reciprocating, rotary screw, rotary vane, scroll, and centrifugal. Within each compressor type, individual design as well as the refrigerant used affects the choice of lubricant. Although generalizations are often made, care must be exercised when selecting lubricants as each system is often unique in at least one aspect.1-4 Reciprocating compressors are utilized in a variety of applications and have a multitude of designs. Principal areas of lubrication include the crankshaft and associated bearings, connecting rods, wrist pins, pistons, cylinders, piston rings, and valves. Lubricant supply to the required area can be accomplished by methods ranging from simple “splash” to more complex pressurized forced-feed systems. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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Rotary screw compressors require lubrication of the rotors, bearings, and shaft seals. Flooded machines of this type require the lubricant to remove heat from and to seal the rotors. Oil is typically injected into the rotors with the refrigerant gas and sees higher refrigerant gas concentrations and pressures than are found in reciprocating systems. Usually oil is supplied to the bearings by pressurized forcedfeed systems. Rotary vane and scroll compressors require lubrication of bearings and shaft seals. Both of these types of compressors have severe lubricant requirement in the “tip” portion of the compressor (rotor tip and/or scroll tip). Oil is supplied by entrainment in the return gas, splash, or by oil pumps. Centrifugal compressors typically have bearings, gear reducers and seals which require lubrication. Generally, lubrication is provided by some type of pumping system. REFRIGERANT AND LUBRICANT FACTORS The wide variety of refrigerants used today (e.g., hydrochlorofluorocarbons (HCFC), hydrofluoro-carbons (HFC), ammonia, hydrocarbons, carbon dioxide, etc.) afford significantly different properties when used with a given lubricant. Miscibility (liquid/liquid interaction) and solubility (gas/liquid interaction) can vary greatly from refrigerant to refrigerant and impact properties such as working viscosity, compressor/system efficiency, and lubricant return to the compressor. Because of the wide temperature extremes found in refrigeration systems, the temperature/viscosity characteristics of the lubricant are extremely important as are the low temperature properties. Also important, particularly with some of the synthetic lubricants, is system cleanliness. Contaminants in the system can lead to a number of potential problems including plugging of expansion devices, copper plating, and lubricant breakdown.

LUBRICANT REQUIREMENTS Factors critical to the success of the lubricant in refrigeration applications include viscosity, solubility, miscibility, chemical and thermal stability, and materials compatibility. VISCOSITY Lubricant-related problems often involve the loss of viscosity, which can result in failure in critical areas of the compressor.5 To determine the proper lubricant viscosity, the first consideration is the viscosity of the lubricant itself. Most lubricants are supplied to an ISO viscosity grade which specifies the viscosity at 40°C. Table 1 shows typical viscosity grades used in various refrigeration systems. The fluid viscosity–temperature relationship (viscosity index, VI) becomes more important with the increased use of synthetic lubricants. Different classes of lubricants provide different viscosity behavior with varying temperature as shown in Table 2. Since the working temperature of a compressor is rarely 40°C, knowledge of the lubricant viscosity at operating temperature is important. Viscosity of the lubricant itself is not the only consideration when selecting the appropriate viscosity grade. The most common cause of reduced viscosity is excessive dilution of the lubricant with refrigerant which results from improper miscibility and/or solubility of the refrigerant/lubricant pair. A liquid refrigerant which mixes in a clear, single phase with the lubricant is miscible; separation into two separate phases reflects immisciblity. These miscibility characteristics depend on temperature and lubricant concentration. As an example, Figure 1 compares the miscibility of various lubricant classes with HCFC-22. Figure 2 compares the miscibility of an ISO 32 POE in R-134a, R-507, and R-407C. The use of synthetic lubricants complicates the miscibility criteria. Even different lubricants within the same class and ISO grade can show vastly different miscibility characteristics, as illustrated in Figure 3 with various ISO 32 polyol ester lubricants and R-134a refrigerant.

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Solubility characteristics of lubricants can vary greatly, depending on the pressure and temperature of the system as well as the chemical composition of the lubricant. Significant solubility can greatly impact the working viscosity of the refrigerant lubricant pair. Figures 4 to 9 show examples of pressure–temperature–viscosity data, used to predict refrigerant dilution for various refrigerant/lubricant pairs. Overdilution of the lubricant with refrigerant can lead to many problems. Occurrence at compressor sump conditions (either during operation or shutdown) can lead to compressor failure due to insufficient lubrication. Excessive refrigerant solubility may also increase the foaming tendencies of the lubricant, which can interfere with lubricant delivery.1-6 The lack of solubility at evaporator conditions can result in the accumulation of lubricant out in the system away from the compressor sump. This can lead to oil starvation of the compressor and ultimate failure. CHEMICAL AND THERMAL STABILITY Lubricant breakdown as a result of poor chemical or thermal stability can potentially lead to problems. One extreme can result in the formation of varnish, sludge, or other insoluble deposits. The opposite extreme can lead to reduced lubricant viscosity. Either situation can result in compressor and/or system-related failures. ASHRAE/ANSI 97-1989 Sealed Tube Stability Test7 is used to evaluate this stability for any potential refrigerant/lubricant pair.

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FIGURE 1 Miscibility of lubricants with R-22.

FIGURE 2 Miscibility of ISO 32 POE with (Lines A and D) R-507; (Line B) R-407c; (Line C) R134a. MATERIALS COMPATIBILITY Polymeric materials of construction (e.g., insulating materials, wire enamels, and elastomers) used in refrigeration systems must be compatible with the refrigerant lubricant pair. Extensive evaluation of materials has been done and is collected in the ARTI Refrigerant Database.8 Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Miscibility of ISO 32 POEs with R-134a.

FIGURE 4 Viscosity-temperature-pressure chart for R-134a and ISO 32 POE.

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FIGURE 5 Viscosity–temperature–pressure chart for R-134a and ISO 32 POE.

LUBRICANT TYPES A wide variety of lubricant types available for refrigeration applications allows the lubricant supplier to tailor lubricants for specific applications. Table 3 gives an overview of the key properties of the lubricants most commonly used. MINERAL OILS Highly refined mineral oils, similar to white oils, have been used for refrigeration applications for many years. These petroleum-base lubricants can vary in their physical properties, chemical structure, and degree of refining — all of which influence performance in refrigeration applications. They are generally classified as either paraffinic or naphthenic. Paraffinic mineral oils generally contain predominantly linear or straight-chain paraffins. These oils are typically available in viscosity grades ranging from ISO VG 10 to ISO VG 100. Generally the higher viscosity grades have poorer low temperature properties. They are currently used in Copyright © 1997 CRC Press, LLC.

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FIGURE 6 Viscosity–temperature–pressure chart for R-134a and ISO 68 polyol ester. (From Short, G. D., CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994.)9 HCFC, ammonia, and hydrocarbon applications. Utilization of these oils in HFC applications has been limited due to their immiscibility with these refrigerants. Naphthenic refrigeration oils contain higher levels of unsaturated aromatic molecules. They have better low temperature properties than the paraffinics and are generally available in ISO VG 10-100 grades. Naphthenics have slightly higher miscibility characteristics than paraffinics and are used in the same applications. These oils are also immiscible with the new HFC refrigerants.

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FIGURE 7 Viscosity–temperature–pressure chart for R-134a and ISO 68 polyalkylene glycol. (From Short, G. D., CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994.)9 SYNTHETIC OILS Synthetic fluids offer a wide range of chemical classes and functionalities. As a result, generalities about these fluids are difficult to make. The use of synthetic oils for refrigeration was first proposed in 1929,10,11 to solve problems with mineral oils such as wax separation, low miscibility with some refrigerants, and carbonization of valves in reciprocating compressors. Additional advantages for Copyright © 1997 CRC Press, LLC.

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FIGURE 8 Viscosity–temperature–pressure chart for HCFC-22 and ISO 32 alkyl benzene. (From Short, G. D., CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994.)9 some of these lubricants may include improved stability in the presence of refrigerants at high temperatures, and better viscosity/temperature characteristics resulting in improved hydrodynamic lubrication and better lubricity in the presence of refrigerants. Each category of lubricant represents a broad class of base fluids which can have greatly different properties. Development of synthetic-based lubricant for refrigeration applications is expected to continue for the next several years. Copyright © 1997 CRC Press, LLC.

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FIGURE 9 Viscosity-temperature-pressure chart for HCFC-22 and ISO 32 naphthenic mineral oil. (From Short, G. D., CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994.)9 Alkyl benzenes (AB) are used in some R-22 and R-12 systems and with ammonia. The aromatic portion of this molecule enhances miscibility with HCFCs, while the alkyl chain is used to build viscosity. This limits the effective use of this lubricant to lower viscosity grades of ISO VG 22-100. Blends of synthetic oils with mineral oils are sometimes used to enhance the properties of the mixture and minimize potential detrimental effects. The most common blends are those of mineral oil and alkyl benzenes. Mineral oil/alkyl benzene blends produce additional miscibility compared Copyright © 1997 CRC Press, LLC.

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to mineral oil alone, while maintaining the better lubricity of the mineral oil. Figure 10 shows the miscibility improvement with an alkyl benzene/mineral oil blend with HCFC-22. Polyalphaolefins (PAO) are available in viscosity grades ranging from ISO VG 5 to 1000. These synthetic hydrocarbons are immiscible with HFC refrigerants and have seen only limited use in these systems. PAOs are used as high-stability ammonia oils. The low temperature properties of these lubricants enables flow at temperatures of and lower.

FIGURE 10 Miscibility of a mineral oil/alkyl benzene blend with R-22. Polyol esters (POE) have seen signifiacant developement for use in refrigeration and are available for all HFC refrigeration applications. High viscosity grades have also found utility in some HCFC Copyright © 1997 CRC Press, LLC.

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systems. This class of compounds affords an extremely wide range of products of different properties and viscosity grades. Polyol esters cannot be used in ammonia applications because they chemically react with the refrigerant. Polyalkylene glycols (PAG) lubricants have found utility in a number of refrigeration applications and are the lubricant of choice in automotive air conditioning with R-134a. They are also finding use in some hydrocarbon and ammonia systems. Like POE, PAGs afford a wide variety of lubricant properties and viscosity grades. ADDITIVES Additives are sometimes used in refrigeration lubricants. The most common types of additives include: stability improvers, lubricity aids, and antifoam agents. Need for additives is dependent on the lubricant, refrigerant, and equipment design. The lubricant qualification procedure must be used to insure there are no detrimental effects with the additives.1

RETROFIT OF EXISTING SYSTEMS One of the most important issues facing the refrigeration industry is that of retrofitting the existing systems which are currently using CFC-based refrigerants. While some of these systems are being replaced by new environmentally friendly units, many existing systems have several years of viable service remaining. For these systems, retrofitting offers a cost-effective way of changing the system to the newer HFC-based refrigerants. The primary problem with mineral oils left in a system which has been converted to an HFC refrigerant is that they are often immiscible with the refrigerant. This can lead to reduced system efficiency by fouling or coating of the evaporator, and reduced bearing life by adversely affecting the lubricity of the fluid. Even low amounts of mineral oil can also reduce the miscibility characteristics of the synthetic lubricant to an unacceptable level. This is illustrated in Figure 11.

FIGURE 11 Effect of mineral oil concentration on miscibility of ISO 68 POE with R134a. To prevent this type of problem from occurring, the mineral oil can be removed by flushing the system with the replacement synthetic lubricant. This can be accomplished by the following guidelines:

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• Transfer the current refrigerant charge to a receiver. This will isolate the refrigerant, keeping it from being lost when the system is opened to drain the mineral oil. • Drain the mineral oil from the compressor. For certain small, hermetic systems without a preinstalled oil drain, this may be accomplished only by cutting the compressor from the piping and tipping the compressor on its side. In larger systems, the many possible areas where the lubricant can collect should also be drained. • Recharge the machine with the new lubricant and reattach the machine. The amount of fluid to be added to the system should not exceed the amount of mineral oil which was drained. Since some mineral oil may be left in the system, installing a full charge of new fluid may overfill the system. • Recharge with the CFC refrigerant and run for a period of time. CFC refrigerants are miscible in mineral oil as well as the new synthetic lubricants. By running the system with the CFC refrigerant, mineral oil in the evaporator and other inaccessible portions of the system is flushed out and returned to the oil sump. The length of time necessary to complete this portion of the operation is dependent on the size of the system and the complexity of the piping involved. A large, extended system, with hundreds of feet of piping will require a much longer flush interval than a smaller, compact system. • After this running period, a sample should be taken to determine the mineral oil level in the lubricant. If the amount of mineral oil is still above the maximum level set by the system manufacturer, the previous steps must be repeated until the residual mineral oil content meets acceptable levels: 1 to 5%, depending on the system design and manufacturer. • Charge the system with the HFC refrigerant and fresh fill of lubricant. Several other areas which require attention during a retrofit include changing the filter dryers and expansion valves. In addition, monitoring of the system performance before and after the retrofit will assist in determining if the system is operating at the expected capacity. The system should also then be reidentified with the correct refrigerant and lubricant.

REFERENCES 1. Spauschus, H.O., Evaluation of lubicants for refrigeration and air conditioning compressors, ASHRAE J., 26, (5), 59, 1984. 2. Kruse, H.H. and Schroeder, M., Fundamentals of lubrication in refrigeration systems and heat pumps, ASHRAE J., 26, (5), 5–9, 1984. 3. Daniel, G., Anderson, M.J., Schmid, W., and Tokumitsu, M., Performance of selected synthetic lubricants in industrial heat pumps, J. Heat Recovery Syst., (2), 4, 359–368, 1982. 4. Short, G.D., Synthetic lubricants and their refrigeration applications, Lubr. Eng., 46, (4), 1990. 5. Burkhardt, J. and Hahne, E., Surface tension of refrigeration oils, IIR — Commisions B1, B2, El, E2, Mons, Belgium, 1980, 111. 6. Reese, L., Accessible hermetic compressors, Warmepumpen-Grundlagen-Komponenten-Auslegung- Bau and Betrieb, Vulkan-Verlag, Essen, Germany, Ed. 1, 1978, 100. 7. ASHRAE Standard 97-1989, American Society of Heating, Refrigerating, and Air Conditioning Engineers, Atlanta, GA, 1986. 8. Calm, J.M., ARTI Refrigerant Database, Air Conditioning and Refrigeration Technology Institute, Arlington, VA, July, 1996. 9. Short, G.D., Refrigeration and Air Conditioning, CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994, 387–408. 10. Sanvordenker, K.S. and Larime, M.W., ASHRAE Trans., 78(2), 1990. 11. Shoemaker, B.H., Symposium, Synthetic Lubricating Oils, Ind. Eng. Chem., 42(12), 2414, 1959.

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35 Space and High Vacuum Lubricants Robert L. Thom and Michael R. Hilton CONTENTS Solid (Dry) Lubricants...................................................................................................................................355 Fluid Lubricants — Oils...............................................................................................................................358 Greases .............................................................................................................................................................360 References.........................................................................................................................................................362

Three types of lubricants are used in vacuum environments: solids (dry), oils, and greases. Descriptions of most solid and fluid space lubricants are given, together with conditions for use, in the NASA handbook by McMurtrey.1 This section will define and review the lubricants available, including an assessment of their favorable and unfavorable properties. Whenever appropriate, methods of application or processing will be reviewed.

SOLID (DRY) LUBRICANTS Four types of solid or dry lubricants are available for vacuum applications: soft metals, lamellar solids, polymers, or other soft solids (see Table 1). Composites of these four types of lubricants or combinations of one or more of them with matrix or support materials are also available. SOFT METALS Soft metals, including lead, gold, silver, and indium, have all been used as lubricants in vacuum applications.2 Of these metals, lead has had the most success and use. To apply lead, burnishing or electroplating has been used; however, deposition by ion plating provides the best adhesion and is preferred for (uniform) coverage. Optimum performance of lead and other metals is achieved at approximately 1 thickness. Ion-plating lead films have been particularly effective in spacecraft bearings found in solar array drive mechanisms, especially in European satellites. Silver and gold are useful in situations requiring electrical conductivity. However, silver is generally too hard for most applications, and gold work-hardens quite easily. Lead remains soft at room temperatures, and evidence indicates that it can lubricate at 20K. LAMELLAR SOLIDS Lamellar solids in relatively wide use as lubricants include the disulfides and diselenides of Mo, W, Nb, and Ta. Graphite is also a lamellar solid lubricant, but the pure material is not suitable for vacuum applications since it depends on water vapor for low-shear qualities. The anisotropic, planar crystal structures of lamellar solids provide low-shear planes for lubrication. These solids also have Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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high load-bearing capacity when compressed in a direction perpendicular to their low-shear planes. This load-bearing capability of the lamellar solids is an advantage over solid polymer lubricants. Of the lamellar solids, MoS2 films deposited by sputtering have been the most widely investigated and developed since early in the space program3 and especially in the last few years.4.5 MoS2films have a lower friction coefficient than Pb films ( 0.01 vs. 0.1 in vacuum, respectively), which lowers mechanism torque and power consumption. MoS2 films are also superior to Pb films in pure sliding applications. Sputter-deposited MoS2 has superior endurance and lower running friction than either burnished or bonded MoS2. An important property of MoS2 is its high compressive strength relative to steel — the lubricant will not cold flow out of highly loaded contacts that remain static until use during flight. The performance of sputter-deposited MoS2 is critically dependent upon film microstructure, which includes composition, morphology, crystallinity, and preferred orientation.6 These properties, in turn, are very dependent upon deposition conditions; the presence of water vapor during deposition is a particularly insidious variable.7 The general trend in film development in recent years has been the production of dense films with low porosity, because porosity leads to large-scale film debris generation early in wear.6 Most films grow with their low-shear basal planes perpendicular to the substrate. Reorientation of the basal planes to a parallel alignment with the substrate occurs during wear. Stressinduced crystallization has also been observed after sliding wear in some dense films that were disordered as-deposited.8 There are several deposition practices that can yield these dense films, including high growth rates,9 low deposition pressures,10 ion bombardment during film growth,11,12 and the incorporation of dopants (Au, Ni, water vapor) that are either co-sputtered continuously,13 or deposited as multilayers.14,15 Some of these films have an initial preferred orientation of low-shear basal planes parallel to the substrate. A recent testing program has extended the known limits of performance of sputter-deposited MoS2 in precision gimbal bearings.16-18 Specifically, endurances of 48 million cycles have been

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achieved for small angular contact bearings appropriate for gimbals. Lifetimes of several million cycles have also been achieved for larger bearings. Metal-MoS2 multilayer films, to date, have the best-controlled fracture toughness properties which affect spallation.19,20 POLYMERS AND POLYMER COMPOSITES Polymers, consisting of anisotropically bonded molecules, can provide low friction surfaces in vacuum, if the molecule chains align properly at the contacting surface. However, the loadbearing capability of polymers is generally low, so additives are required to strengthen the polymer to avoid ploughing into the bulk. For vacuum applications, polymer composites rather than pure polymers are generally used.21,22 Since these composites have structural integrity, selflubricating composite components can be fabricated that can, in principle, provide a continual source of solid lubricant to critical components. To date, polytetrafluoroethylene (PTFE) has been the polymer used the most in vacuum. This is because PTFE performs well in vacuum and in the presence of absorbed vapors. However, PTFE has a tendency to cold-flow under load, necessitating a binder to restrain the polymer bulk, i.e., to prevent ploughing. Some polyimides appear to be excellent in vacuum because they exhibit low friction coefficients without significant cold-flowing of the bulk.23 However, polyimides are very sensitive to water vapor absorption. Water molecules appear to hydrogen bond to the polymer molecules and then inhibit molecular shear. Thermal pretreatment of polyimides appears to be essential for good performance in vacuum. When making polymer composites, other materials are added to the polymers for several reasons: to increase load-carrying capacity, to lower the friction coefficient, to promote a low wear rate, and to increase the composite’s thermal conductivity. Fibers and particulate additives can be used, although fibers are more effective for increasing composite load-carrying capacity. Studies indicate that MoS2 in some composites facilitates polymer transfer to a critical component; the polymer is the primary lubricant, not the MoS2.24 Table 2 lists some self-lubricating composites that are available and their possible uses in space. When a polymer composite is selected for an application, contact stress is probably the most important property to consider. The ideal composite must support, in the bulk of the material, the dynamic stresses of the application and must allow for the formation, by local deformation, of a low-shear layer at the surface.

An important application of polymer composites is in the fabrication of bearing retainers. For solid lubricated bearings, a variety of materials are now available, as noted in Table 2. These retainers can be made with solid lubricants as a constituent. Such sacrificial or active retainers provide, or augment, lubrication to the bearing raceways by material transfer. Such transfer is difficult to control Copyright © 1997 CRC Press, LLC.

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uniformly — torque noise can result from uneven transfer. For torque noise-sensitive applications, retainers that are not sacrificial and that operate in a benign manner with solid lubricant films on the raceways are needed. Such benign retainers need further development. By far the largest application of polymer composite retainers are for fluid lubricated bearings. Retainers consisting of cotton fiber-reinforced phenolic resin have the most use. A polyimide composite retainer material has been developed that has had some use in space systems.25 OTHER (NONPOLYMER-BASED) COMPOSITES Two examples of nonpolymer-based composites are particularly worthy of attention. One type, leaded bronze composites, have been fabricated into bearing retainers. When used in conjunction with a lead-coated bearing (for example, in solar array drive mechanisms), the lead in the retainers provides an effective supplemental source of lead when the original film is worn. In another example, composite blocks of silver, MoS2, and either graphite26 or copper27 are used as brushes in sliding electrical contacts. The silver provides conductivity and structural integrity, the MoS2 lubricates in vacuum, and the graphite or copper may lubricate in air.

FLUID LUBRICANTS — OILS There are several types or categories of liquid lubricants that have been used or could be used for vacuum/space applications. These categories include: (1) silicone oils; (2) mineral oils; (3) perfluoropolyalkylethers (PFPEs); and (4) other new synthetics (including multiply-alkylated cyclopentanes [MACs], poly- -olephins [PAOs], and polyolesters [POEs]. Except for gyroscope applications, these lubricants generally encounter boundary contact conditions at least at some time during their service life. SILICONE OILS The low vapor pressures and low pour points of some silicone oils led to their early use in space applications. However, these oils are only moderately effective lubricants. One problem is that some of these oils tend to form polymers on the bearing surface, which leads to torque noise. Another problem is that these oils creep readily on metal surfaces. Because of these problems and the availability of better alternatives, silicone oils would not be used on contemporary spacecraft. However, these oils are used as damper fluids and as thermal conduction media in some instances. MINERAL OILS Highly refined mineral oils have been a popular choice for sealed mechanisms, such as momentum wheels, reaction wheels, and despin mechanisms. A series of super-refined gyroscope lubricants (SRG and KG-80) have been available and comprise a homologous group of natural polymers that allows the designer to choose a fluid having particular viscosity characteristics for a specific application.28 Mineral oils also can be formulated with antiwear and other additives. Due to their high vapor pressures and wide distribution of molecular weights, these natural oils are being replaced by synthetic oils in space applications. PERFLUOROPOLYALKYLETHERS (PFPEs) PFPE lubricants have lower vapor pressures, lower pour points, and higher viscosity indexes than mineral oils (see Table 3). Thus, they are useful in space mechanisms that are not completely sealed or that are somewhat cooler (200K) than would be acceptable for mineral oils. In particular, one of the PFPEs (Fomblin Z25) has a very high viscosity index and is exceptionally useful over a wide temperature range.

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PFPEs perform reasonably well under nonboundary contact conditions. Conventional antiwear additives do not dissolve into the PFPE fluids, although a new class of compatible additives has been reported.29 Lubricant degradation by polymerization leads to increased bearing torque noise and wear. The substrate-induced degradation can be retarded by substituting one or both of the steel surfaces with either ceramic components or ceramic-coated steel (or presumably by using the new additives). TiC- and TiN-coated steel and Si3N4 components have shown improved performance. PFPEs have extremely low surface tensions (18 dyne/cm) and, therefore, creep very readily over metal and other surfaces. Because of their similar chemical structures, the lubricants also dissolve fluorocarbon coatings that are used as antimigration barriers. The available commercial PFPEs and their properties are listed in Table 3. OTHER SYNTHETIC LUBRICANTS MAC, PAO, POE, and other polymer oils can be synthesized and blended to produce viscosity, vapor pressure, pour point, and other properties in a controlled way that will suit various needs.26,30,31Vapor pressures that are almost as low as those of linear PFPEs have been obtained for the MAC oils.32 The vapor pressures of PAOs and POEs are not as low, but they can be lower than those of conventional mineral oils, as noted in Table 3. A variety of laboratory testing has shown that the synthetic hydrocarbons and esters have longer endurance, due both to the lower volatility of these oils and to their better chemical stability. Laboratory screening tests have shown that synthetic hydrocarbons give the longest wear lifetimes in a simulated boundary-lubrication test facility. Bearing tests with a fixture designed to simulate the oscillatory motion of a weather scanner have shown that a PAO provides near freezing (0°C) temperature capability and significantly outlasts both a silicone oil and a PFPE.33 PAO oils have given very good performance in lightly loaded, high-speed gyroscope bearings. Copyright © 1997 CRC Press, LLC.

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Another relatively new class of synthetic lubricants with vapor pressures acceptable for vacuum applications and the capability to be compounded with additives is known by the term “silahydrocarbons.”34 Their tribological performance has not been thoroughly tested in specific applications, but the results of conventional four-ball and traction tests are very encouraging.35

GREASES Greases are comprised of oils compounded with a pore network-forming thickener, such as a soap or a fine particle suspension. For results of an extensive testing program of greases in vacuum, see the reference by McMurtrey.36 Depending on the type of oil and the nature of the thickener, greases can be formulated for various applications involving a variety of components with different types of contact (e.g., slow or high-speed angular contact ball bearings, journals, gears). Oils in greases can be from any of the categories discussed in previous sections. However, the solubility properties (i.e., chemical compatibility) of the oil determine the selection of thickener and, hence, the grease properties. Mineral oils and certain synthetics have good solvent properties, so they can be formulated with soaps of different cations to make what are known as channeling greases. Such greases are pushed out of the way and form a path (channel) when the balls of a bearing pass through the grease. When working properly, oil will continually diffuse out of the mounds of grease on the edges of the ball path to supply lubricant to the contacting surfaces. If a grease is fluid enough that it tends to fill the spaces between balls, it is a “slumping” (nonchanneling) grease. The consistency of a grease depends on the type of thickener used and the relative amounts of oil vs. thickener. The primary role of grease in a vacuum application is as a reservoir for supplying oil to contacting surfaces. A bearing properly packed with grease will also suffer less oil loss by creep or physical spattering because of the physical barrier the grease can provide. However, the lubrication properties of any grease can be only as good as those of the base oil, so care must be exercised in selection of the base oil. For example, formulation of a volatile oil into a grease cannot prevent the oil from contaminating a vacuum system; a low-volatility oil must be used. Real time long-term tests were performed at the Marshall Space Flight Center to determine adequate lubricants for use in space environments. Select data are provided from McMurtrey’s report of long-term testing performed in a vacuum environment under several different conditions. These conditions included a constant temperature of 38°C and 93.3°C. A start-stop test was also performed with the motors held at ambient lab temperature. All three of these conditions were performed for two lengths of time, 1 year and 5 years. The tests were continuous for those periods of time, i.e., 24 hours a day, 7 days a week until either failure of the lubricants occurred or the time period was reached. Over 30 different grease samples were evaluated in each of the conditions for the 1-year duration test. Data of the top 10 performers are provided for these tests and the 5-year tests include data from four different greases. Table 4 exhibits data from greases tested at 38°C. All of the top ten greases made it through the 1-year test with no failures. They are sorted in order of average weight loss. The 5-year test was better able to discriminate between the four tested greases. Even so, the top two greases had no failures for the 5-year period of time. Table 5 exhibits data from the same group of 35 greases tested at a temperature of 93.3°C. Again, there are no failures of the top ten greases tested for 1 year. However, after 5 years, none of the greases made it through the test period unscathed. Also, at the higher operating temperature the average mass loss is significantly greater than for greases evaluated at 38°C. Table 6 exhibits data from the start-stop test performed. This test increased the severity of the evaluation by forcing a boundary lubrication regime into the test. As in the previous two tables, there are no failures for the 1-year test. However, after 5-years, only one lubricant was able to make it thorough the entire test with no failures. This test data represents performance under strict laboratory conditions which may or may not be similar to the design used in other vacuum/space applications. For this reason, the data should be used judiciously. Copyright © 1997 CRC Press, LLC.

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REFERENCES 1. McMurtrey, E. L., Lubrication Handbook for the Space Industry — Part A: Solid Lubricants; Part B: Liquid Lubricants, NASA TM-86556, NASA George C. Marshall Space Flight Center, Huntsville, AL, Dec. 1985. 2. Roberts, E. W., Tribol. Int., 23(2), 95, 1990. 3. (a) Spalvins, T., ASLE Trans., 14, 267, 1971; (b) Ibid., 17, 1, 1973. 4. See, for example, articles in New Materials Approaches to Tribology: Theory and Applications, Pope, L. E., Fehrenbacher, L. L., and Winer, W. O., Eds., Mat. Res. Soc. Sump. Proc, 140, 215–284, 1989. 5. See, for example, articles appearing in Thin Solid Films: 154, 309–332, 1987; 181, 461–495, 1989; and in Surf. Coat. Technol, 36, 329–359, 1988; 43/44, 622–629, 1990. 6. Hilton, M. R., Bauer, R., and Fleischauer, P. D., Thin Solid Films, 188, 219, 1990, and references therein. 7. Buck, V., Thin Solid Films, 139, 157, 1986. 8. Hilton, M. R. and Fleischauer, P. D., Mat. Res. Soc. Symp. Proc., 140, 227, 1989. 9. Nabot, J.-Ph., Aubert, A., Gillet, R., and Renaux, Ph., Sur. Coat. Technol, 43/44, 629, 1990. 10. Muller, C., Menoud, C., Maillat, M., and Hintermann, H. E., Surf. Coat. Technol., 36(1-2), 351,1988. 11. Kuwano, H. and Nagai, K., J. Vac. Sci. Technol. A, 4, 2993, 1986. 12. Bolster, R. N., Singer, I. L., Wegand, J. C., Fayeulle, S., and Gossett, C. R., Surf. Coat. Technol., 46, 207, 1991. 13. References on porous films with less than 5% nickel: (a) Stupp, B. C, Thin Solid Films, 84, 257, 1981; (b) Stupp, B. C, Proc. Third Int. Conf. Solid Lubrication, ASLE SP-14, 7-10 August 1984, Denver, CO, STLE, Park Ridge, IL, 217–222. 14. Hilton, M. R., Bauer, R., Didziulis, S. V., Dugger, M. T., Keem, J., and Scholhamer, J., Surf. Coat. Technol., 53, 13, 1992. 15. Jayaram, G., Marks, L. D., and Hilton, M. R., Nanostructure of Au–20%Pd Layers in MoS2 Multilayer Solid Lubricant Coatings, Surf. Coat. Technol., 76–17, 393, 1995. 16. Loewenthal, S. H., SDIO Tribomaterials/Precision Gimbal Demonstration Program Phase 2 Final Report, Wright Laboratory WL-TR-94-4094, Dayton, OH, December 1993.

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17. Loewenthal, S. H., Chou, R. G., Hopple, G. B., and Wenger, W. L., STLE J., 162–164, 919, 1993. 18. Hopple, G. B. and Loewenthal, S. H., Surf. Coat. Technol., 68/69, 398, 1994. 19. Hopple, G. B., Keem, J. E., and Loewenthal, S. H., Wear, 162–164, 919, 1993. 20. Hilton, M. R., Surf. Coat. Technol., 68/69, 407, 1994. 21. Gardos, M. N., Self-lubricating composites for extreme lubricating conditions, in Friction and Wear of Polymer Composites, Klaus, F., Ed., Elsevier, Amsterdam, 1986, 397–447. 22. Fusaro, R. L., Tribol. Int., 23(2), 105, 1990. 23. Fusaro, R. L., Tribol. Trans., 31(2), 174, 1988. 24. Sutor, P. and Gardos, M. N., in Proc. Third Int. Conf. Solid Lubrication, ASLE SP-I4 (7–10 August 1984, Denver, CO), STLE, Park Ridge, IL, 1984, 258. See also Ref. 32, pp.154–163. 25. Meldin 9000. 26. Fleischauer, P. D. and Hilton, M. R., Mat. Res. Soc. Symp. Proc., 140, 9, 1989. 27a. Roberts, E. W., A Review of Sliding Electrical Contacts for Space Applications, European Space Agency/European Space Tribology Laboratory (Risley, U.K.), ESA Report, Contract No. ESA 4099/79/NL/PP, ESA(ESTL)52, (October 1981). 27b. Roberts, E. W., Sliding electric contacts in space: observations on existing technology and new trends in low-speed applications, in Proc. First European Space Mechanisms and Tribology Symposium, (October 1883, Neuchatel, Switzerland) ESA SP-196, 3-10. 28. Kannel, J. W. and Dufrane, K. F., Rolling element bearings in space, Twentieth Aerospace Mechanisms Symposium, NASA CP-2423, 1986, 125–132. 29. Sharma, S. K., Gschwender, L. J., and Synder, C. E., Jr., J. Syn. hub., 7(1), 15, 1990. 30. See, for example: (a) Mobil Chemical Tech. Bull., SHFBSD72, Mobil SHF Base Stocks (discusses Polyalphaolefins). (b) Nye Specialty Lubricants Tech. Bull., Synthetic Hydrocarbon Precision Lubricating Oils (discusses Polyalphaolephins). (c) Bray Oil Co. Tech. Bull, NPT 4 Lubricating Oil, Synthetic for Low and High Temperatures (discusses synthetic ester base fluids), (d) Pennzoil Products Co. Tech. Bull. No. CF0BF1 (February 1991), Pennzane Synthesized Hydrocarbon Fluid X 2000 (discusses multiply alkylated cyclopentanes). 31. Venier, C. L. and Casserly, E. W., Multiply-alkylated cyclopentanes (MACs) — A new class of synthesized hydrocarbon fluids, Lub. Eng., 47(7), 586, 1991. 32. Venier, C. L., (Pennzoil Products Co.), private communication to M. R. Hilton (Aerospace Corp.), The Woodlands, TX, May 1991. 33. Carré, D. J., Fleischauer, P. D., Kalogeras, C. G., and Marten, H. D., J. Tribol, 113, 308, 1991. 34. Snyder, C. E., Jr., Tamborski, C, Gschwender, L. J., Chen, C. J., and Anderson, D. R., ASLE Trans., 25(3), 299, 1982. 35. Sharma, S. K., Snyder, C. E., Jr., and Gschwender, L. J., Tribological properties of some advanced space lubricants, Tribol. Trans., in press. 36. McMurtrey, E. L., High Performance Liquid and Solid Lubricants — An Industrial Guide, Noyes Data Corp., Park Ridge, NJ, 1987.

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Steel Industry Lubricant Guidelines Richard C. Schrama and Daniel D. McCoy

INTRODUCTION These steel industry lubricant guidelines have been developed to supply maintenance and lubrication personnel in the steel mills and the lubricant manufacturers alike with some performance requirements for various lubricants. These requirements or lubricant properties are benchmarks, not specifications to be followed as standards in the industry. The lubricant types covered are: circulating oils, hydraulic oils, gear oils, greases, fire-resistant lubricants, and speciality lubricants. The performance guidelines or properties for the lubricants are based on standard ASTM and industry-accepted test methods. There is no attempt made to provide guidelines for rolling oils used in the cold rolling and tempering of strip and for oils used in the hot rolling of steel, since the properties of the oil and the formulations are tied into the performance of the mill and rolls. These oils can be classified as metalworking fluids and treated in the manner as cutting or forming fluids. In most cases rolling oils are made up of animal fats or synthetic esters, biocides, surfactants, and pH modifiers. They are then mixed with water to create a solution of between 0.5% to 5% concentration. The rolling oil performance is tied to the surface finish of the steel, heat transfer rate to remove the frictional heat of rolling, and cleanliness of the strip. This chapter also does not deal with the other process fluids used in the steel mill. These include: pickling oils, preservative oils and coatings, coating oils for the prevention of rust on the steel surface, cleaning solvents and agents, and general maintenance oils and coatings. These lubricants would be used for most steel industry applications including: pumps, steam turbines, gearboxes, pinion gearboxes on rolling mills, oil film bearings for backup rolls, slewing rings, caster turrets, caster segments, mobile equipment, cranes, fans, rolling mill work roll assemblies, machine tools, blast furnace valves, blast furnace skip hoisting machinery, and mandrels. These guidelines cover only the minimal lubricant requirements. In many applications, premium products with properties which excede the guidelines are required to meet the application and/or health and safety, disposal, and federal or state regulations. The tables for the lubricant types are: A. Circulating oils Table 1. Turbine Oil Table 3. Circulating Oil Table 4. Extra-Duty Circulating Oil Table 6. Circulating Turbine Oil Table 14.Pale Paraffin Slushing Oil B. Hydraulic oils Table 2. Hydraulic - Extra-Duty, Antiwear Table 10.Inhibited Hydraulic Oil Table 13.Inverted Emulsions Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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Table 15.Noninhibited Hydraulic Oil Table 20.Combination Synthetic–Petroleum Fluids Table 24.Water–Glycol Fluids Table 25.Phosphate Ester Fluids C. Engine oil Table 7. Engine Oil Table 8. Circulating Engine Oil D. Greases Table 5. Ball and Roller Bearing Grease Table 26.Mill Utility Grease Table 27.High Temperature Grease Table 28.High Temperature EP Greases — Complex Soaps and Nonsoaps Table 29.Molybdenum Disulfide Greases Table 30.Roll Neck Grease Table 31.Journal Roller Bearing Grease — AAR Specification M-942-92 E. Rolling oils Table 9. Sendzimir Rolling Oil F. Open gear lubricants Table 11.Open Gear Lubricants — EP Type Table 12.Open Gear Lubricants — Non-EP Type G. Gear oils Table 16.Roll Neck Spray or Gear Oil Table 17.Mist EP Gear Oil Table 18.Extra-Duty Gear Oil Table 19.Hypoid Gear Oil Table 21.Insulating Oil for Transformers and Oil Switches Table 22.Extreme Pressure Oil Table 23.Heavy-Duty Brake Fluid — SAE-J-1703

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REFERENCES 1. Lubrication Engineers Manual, Section 5: Lubricant Performance Guidelines, U.S. Steel Corp. and Association of Iron and Steel Engineers, Pittsburgh, PA, 1971. 2. ASTM Test Methods for Petroleum Fluids, American Society for Testing of Materials, Philadelphia, PA, 1996.

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IV Lubricant Application Systems

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Centralized Lubrication for 37 Industrial Machines James H. Simpson, III* CONTENTS Introduction....................................................................................................................................385 Single Line Parallel Systems.........................................................................................................386 Progressive Feeder Systems..........................................................................................................390 Dual Line Parallel Systems...........................................................................................................390 Pump to Point Systems.................................................................................................................391 Oil and Air Systems.......................................................................................................................391 Planning, Installation, and Maintenance....................................................................................392 System Capacity..............................................................................................................................394 References.......................................................................................................................................395

INTRODUCTION The basic purpose of centralized lubrication is to supply either individual or groups of lubrication points with varying lubricant quantities, from one central position, with the required, metered lubricant quantities. The lubricants may be oil, fluid grease, or grease of NLGI grades 000–2 being applied to machine tools and to a variety of machinery used in applications such as printing, paper processing, metal and plastic production, woodworking, and textile manufacturing. Figure 1 illustrates a basic layout with typical devics for a centralized lubrication system with no return of oil from the lubrication points to the reservoir (total loss design). In a centralized lubrication system (Figure 1), oil is supplied by manually, mechanically, hydraulically, or pneumatically operated piston pumps or intermittently operated gear pumps. Metered quantities of the lubricant, in the range of 0.01 to 1.5 cm3 per lubrication cycle and lubrication point, are dispensed by piston distributors mounted in the tubing system. This quantity delivered to the lubrication points is set by metering nipples fitted to the distributors and by the frequency of lubrication cycles. Of the available types of systems indicated in Figure 2 and compared in Table 1, those used for most applications include: single line parallel, single line progressive, and dual line parallel. Larger systems are broken down into zones to simplify and provide ease of maintenance. These types of systems include: dual line/single line, zoned single line progressive, and zoned single line parallel. Descriptions of the more common system types follow. Automatic lubrication systems are recommended on equipment with ≥ 6 lubrication points or with lubrication application intervals ≤ 200 hours. Manual lubrication systems are recommended on equipment with ≤ 5 lubrication points or with lubrication application intervals ≥ 200 hours.1 Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 Basic system layout of a total loss system. (1) Pump (manually, mechanically, hydraulically, or pneumatically operated piston pump, electrically operated gear pump); (2) main line (connection pump –distributor); (3) piston distributors; (4) secondary line (connection distributor – lubrication point); (5) lubrication point (connection to the friction point of the machine). Automatically controlled systems are additionally provided with the following: (6) Control and monitoring equipment, time- or counter-controlled switch gear, and, if necessary: (7) pressure switch; (8) float switch; and (9) indicator lights.

SINGLE LINE PARALLEL SYSTEMS This type of system is designed using piston distributors, injectors, or restrictors as metering devices. A system pump is used to provide lubricant to each metering device through a piping network. These devices are actuated at the same time and distribute the desired metering quantity to their lubrication point. The quantity of lubricant delivered is determined by the metering device’s allowable amount per lubrication cycle (nipple) and the frequency of such cycles. PISTON DISTRIBUTORS The basic concept of a piston distributor is a two-cycle effort: the fluid used to propel the piston will be the fluid metered on the subsequent stroke. System designs using piston distribution have to provide the following fluid cycle:

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FIGURE 2 Types of centralized lubrication systems. 1. Pressurize the main line: during this cycle the piston distributor meters its fluid to the lubrication points. 2. Depressurize the system: after main line pressure has equaled system pressure and has been maintained for some seconds, the depressurizing cycle is initiated. It returns the pistons home, ready for the next cycle.

To control the system, a pressure switch typically feeds back the information on the system pressure to ascertain that the pressure relief cycle was present during the pause time. The use of piston distributors in centralized lubrication is fairly simple, since all the lube points are parallel and the expansion of a system is facilitated by adding additional metering devices; eliminating lube points is easily accomplished as well, by removing or plugging metering devices (Figure 3). Due to the parallel distribution, each metering device is independent of the others. As long as the main line reaches system pressure, all metering pistons should allocate their metered lubricant quantity into the bearing to which they are connected, unless they encounter back pressure equal to or higher than the system pressure. While piston distributors are produced in many configurations, the standard is the manifold system shown in Figure 3. Piston distributors measure and distribute oil intermittently by means of a pump. The quantities of each lubrication point are defined by the exchangeable metering nipples or metering units. The total amount of oil required can then be regulated by the lubrication frequency. INJECTION OILERS These pneumatically operated piston pumps supply the lubricant in small metered quantities to every lubrication point. Injection oilers can function as group oilers or as single operation distributors. (See Figure 4.) The main header piping size must be chosen so that the back pressure does not exceed 50% of the system pressure. A control method should be implemented to monitor the change in pressure.1

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FIGURE 3 Piston distributor function. (1) The oil allocated for the lubrication point is in front of the piston in the distributor. (2) When the central lubrication pump begins feeding oil, the piston is moved so that the oil in front is delivered by the main line pressure (10–45 bar) to the lubrication point. (3) On release of main line pressure, the piston in the distributor returns to its original position, allowing the oil to flow into the space in front of it.

FIGURE 4 Injection oiler system RESTRICTORS Specific requirements for the restictors (commonly orifices and capillaries) in single line parallel systems are: (1) that the back pressure of the lube points shall not exceed the change in pressure requirements to establish rated flow, (2) the control method should provide a means to measure flow, pressure, and time, and (3) the viscosity range should be limited.

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PROGRESSIVE FEEDER SYSTEMS Progressive feeder systems consist of a pump, progressive feeder blocks (master feeders and secondary feeder blocks), tubing for lubricant distribution, and a control unit. The lubricant is distributed progressively as shown in Figure 5. The sequence of delivery is dependent on the number of sections and the metering quantity of the progressive feeder block. The quantity is determined by the diameter of the spool and the travel within the feeder block. Every spool can only discharge its lubricant if the preceeding outlet has cycled its lubricant. A cycle switch can be used to monitor the feeder’s activity. When the switch is activated, the cycle is complete. The progressive feeder system requires a higher operating pressure than does a piston distributor. Changing the outlet quantity of lubrication for a progressive feeder system is complex, in that the proportional relationship for the entire system has to be recalculated.

FIGURE 5 Progressive feeder systems. (A) Piston side 4 is under pump pressure and piston side 1 has delivered lubricant to outlet 1a. By the movement of piston 1/4, the connection main line — piston side 5 — is free. (B) Piston side 5 is filled with a measured amount of lubricant, and piston 2 will deliver lubricant through outlet port 2a. Next, piston side 6 will be refilled — and so forth — in sequence.

DUAL LINE PARALLEL SYSTEMS In the dual line system, the lubricant is supplied by a pump to the metering devices connected with a two-line header system. The amount of lubricant dispensed can be controlled by varying the size of the metering stroke, altering the metering devices, or changing the frequency of the lubrication cycles.1 Dual line systems can be arranged as one of the following types. Loop type system. The flow charges the header, cycling all metering valves in one direction. After all valves have cycled, the pump flow returns, hydraulically actuating a reversing valve (signaling half cycle). Pump flow is then redirected to the alternate supply line, cycling all the valves in the opposite direction. When the pump flow returns, it resets the reversing valve (signaling full cycle). Nonreturn type system. The flow charges the header, cycling all metering valves in one direction. After all valves have cycled, the pressure rise hydraulically actuates a reversing valve (signaling half cycle). The pump flow is then redirected to the alternate supply line, cycling all the valves in the opposite direction. When the header has filled and built up predetermined pressure, the reversing valve hydraulically resets (signaling full cycle). Copyright © 1997 CRC Press, LLC.

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FIGURE 6 Dual-line feeder operation diagram End-of-line type system. The flow charges the header, cycling all metering valves in one direction. After all valves have cycled, the pressure rise actuates a pressure switch located at the end of the line, which controls a reversing valve (signaling half cycle). Pump flow is then redirected to the alternate supply line, cycling all the valves in the opposite direction. When the pressure rises, a pressure switch located at the end of the alternate line resets the reversing valve (signaling full cycle)1 (see Figure 6).

PUMP TO POINT SYSTEMS With this type of system, oil is supplied to the lubrication points directly with only one feed circuit per point provided in the pump as illustrated in Figure 7.1 This type of system uses multicircuit pumps, cam-actuated piston pumps, or grouped pneumatic/hydraulic piston pumps.

FIGURE 7 Multicircuit pump diagram

OIL AND AIR SYSTEMS This type of system makes use of a continuous air flow carrying a small volume of oil (Figure 8). A mixing device is used to combine air and oil in a designated conductor. The dynamics and friction forces of the air stream are used to break up and move particles through the conductor as evenly spaced droplets along the conductor wall and onto the bearing, or the mixing may occur at the Copyright © 1997 CRC Press, LLC.

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FIGURE 8 Oil + air system diagram.

termination point.1 The metering can be accomplished by using any of the previous systems, but is restricted to the use of oil only.

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PLANNING, INSTALLATION, AND MAINTENANCE The essential points in planning, installation, and maintaining a centralized lubrication system are: 1. Ascertain the number of lubrication points. 2. Estimate the oil quantity required per lubrication point (see Tables 2 and 3 for established formulas for oil) and the total oil requirement per stroke (for piston pumps) or per work cycle (for gear pumps). 3. Select distributors according to metering range and space available. 4. Select pumps according to type of actuation and system capacity. 5. Define type of control for automatic systems (timer- or counter-controlled) and the monitoring, if necessary. 6. In the layout of a system, arrange the run of the main lines and the positions of the distributors in such a way that any air that has entered the system may escape via the lubrication points. For this purpose, mount the distributors at suitable positions and at the end of the system with the outlets to the lubrication points directed upwards. Ensure that the run in the main line rises from the pump to the distributor. 7. When installing especially large and widely branched systems, or if high viscous oils are used, the resistances of the main lines are to be checked, especially for the release process. 8. Confirm the stability of all considerations by an experimental installation, if pure computations will not achieve your goal. 9. Connect each distributor outlet with one lubrication point only. 10. Tighten the bushings of the solderless tube connections firmly, but do not overtighten. (No definite resistance can be felt; a tubing sleeve and tube will be slightly deformed when tightened.)

11. Only connect the secondary line (connection distributor - lubrication point) to the lubrication point after bubble-free oil can be seen leaving the tube end after the pump has been actuated several times. • Longer secondary lines should be filled with oil beforehand. Plastic parts and sealing elements (oil seal rings, lip seals, O-rings) fitted in centralized lubrication units must not be allowed to touch aggressive liquids such as carbon tetrachloride, trichlo-roethylene, aromatic solvents; alkaline solutions and acids. When relatively large oil quantities (not only for tribology purposes) are also utilized for heat dissipation from the lubrication point, a continuous flow of oil should be supplied by gear, Gearotor, or vane pumps. The lubricant for the lubrication points may be optionally distributed via restrictor tubes, metering valves, adjustable metering valves, adjustable metering valve distributors, flow control valves, progressive feeders, or multicircuit gear pumps, from which up to 20 pipelines may be fed directly (or via flow volume dividers) to the individual lubrication points. When oil is distributed via restrictors or multicircuit gear pumps, every lubrication point is continuously fed oil flow with a previously adjusted oil quantity. Progressive feeders, however, are operated by pulsating the flow. Circulating lubrication systems must have an oil return from the lubrication point to the oil reservoir; the returning oil needs to be filtered.

SYSTEM CAPACITY Special consideration during designing is necessary to ensure that the distribution system fluid consumption does not exceed two thirds of the output per stroke or work cycle of a pump. Apply the following to determine the system capacity: Total output of all distributors in system + 25% of this value + allowance of 1 cm3 per m of flexible main line (expansion loss).

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REFERENCES 1. Spiekerman, R., J1751 — Lubrication components and systems used on machine tools and equipment for the automotive industry, SAE, J1751, 1995. 2. Shigley, J.E. and Mischke, C. R., Mechanical Engineering Design, 5th ed., McGraw-Hill, New York, 1989, 480–523. 3. Avollone, E.A. and Baumeister T., III, Mark’s Standard Handbook for Mechanical Engineers, 9th ed., McGraw-Hill, New York, 1987, 197–206. 4. Beitz, W. and Küttner, K.-H., Eds., Dubbel Handbook of Mechanical Engineering, (English edition, edited by Davies, B.J.; translation by Shields, M.J.) Springer-Verlag, London, 1994, Sec. D, Materials Technology, Tribology, pp. D67-D76; and Sec. F, Mechanical Machine Components, Lubrication of Roller Bearings. 5. Centralized Lubrication for Industry, Vogel Centralized Lubrication (Catalog), Pr. 0105 e, EOOO/93. 6. Tribology Centralized Lubrication Systems, Vogel Centralized Lubrication (Handbook), Pr. 0111US, W000/96.

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38 Oil Mist Lubrication Stanley C. Reiber CONTENTS Introduction....................................................................................................................................396 Operating Principle........................................................................................................................396 Application......................................................................................................................................398 Oil Selection....................................................................................................................................399 Temperature Control Requirements...........................................................................................402 References.......................................................................................................................................403

INTRODUCTION Oil mist lubrication was developed in the 1930s by a European bearing manufacturer. Today, few machine elements cannot be lubricated by an automatic, centralized mist system using compressed gas to continuously convey oil mist. In addition to improvements in safety, productivity, and housekeeping, continuous delivery eliminates the over-lubrication that results in periodic applications of lubricant. Lubricant consumption is usually reduced, sometimes by as much as 80’, energy-wasting churning in oil sumps is eliminated, and bearing temperatures are often lowered dramatically.

OPERATING PRINCIPLE The heart of the system is a mist generator (Figure 1), where compressed air, passing through a venturi or vortex, draws oil up into a high-velocity air stream to produce very small oil droplets. After baffles drop larger oil particles back into the reservoir, small particles in the discharge mist generally have diameters less than 6 to 7 µm. With this size, surface tension is high compared to the mass, and droplets wet a surface only on impact at relatively high velocities — hence, the term “dry mist.” This mist is distributed for distances up to 1000 feet or more through pipes, tubing, and hoses sized generally for flow velocities of 15 to 20 ft/s. When the mist flow becomes turbulent above 24 ft/s, the mist particles will strike walls hard enough to stick or “wet out.” Mist flow is metered to each lubrication point by a mist, spray, or condensing fitting with arrangements such as shown in Figures 2 to 5. Since mist fittings only meter flow, they depend on turbulence in and around rolling element bearings to cause oil to wet out directly on bearing elements. Spray and condensing fittings not only control flow but also “reclassify” the “dry mist” into larger droplets. Spray fittings expel oil as fine, wet sprays. Condensing fittings reclassify the oil from a dry mist into coarser, wet sprays or larger drops which then run down adjacent surfaces. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 2 Mist lubrication of rolling element bearings.

APPLICATION Normally, the mist generator output and manifold pressures are commonly 20 inches of water column. Recently, higher pressures have been used for more efficient mist reclassification at delivery points and reduced venting of stray mist to the atmosphere. Mist application is generally based on a standard oil/air ratio in the mist of about 0.4 in.3/h of oil per standard cubic foot per minute (cfm) of mist flow. Table 1 gives a summary of calculations for cfm inputs to various machine elements. These formulas for mist flow requirements may vary somewhat from one manufacturer to another, Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Mist lubrication of plain bearings. depending on their mist generator characteristics. Manufacturers and machinery builders should be consulted for unusual or demanding circumstances.

OIL SELECTION As a rule, oil suppliers should be consulted in selecting appropriate oils for mist systems. While primary consideration is the lubrication requirements of the machine elements, the mist system places some restrictions. Since the oil must be capable of being misted, vigorous foam suppressers,

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FIGURE 4 Mist lubrication of gears.

FIGURE 5 Mist lubrication of chains. Copyright © 1997 CRC Press, LLC.

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tackiness additives, and soap filler should be avoided. Automobile engine oils should not be used because their mistability varies widely, even from lot to lot. Most oils with viscosities of up to more than ISO viscosity grade 1000 can be used when compounded with rust and oxidation inhibitors, EP additives, detergents, and dispersants. Many oil suppliers offer “mist oils” that are slightly mist-inhibited to reduce problems with stray mist.

TEMPERATURE CONTROL REQUIREMENTS Heater recommendations are given in Figure 6 for low ambient temperatures and high oil viscosity. In general, whenever air heaters are used, oil reservoir heaters are also used. Even though heaters might not be required, they are often used to provide a more stable oil/air ratio under widely varying ambient temperatures.

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FIGURE 6 Heater recommendations.

REFERENCES 1. Bornarth, D.M., Oil mist lubrication, in CRC Handbook of Lubrication and Tribology, Vol. 3, CRC Press, Boca Raton, FL, 1994, 409-422. 2. Schrama, R.C., Oil mist vs. air-oil for consumable lubrication systems, Lub. Eng., 49, 8-17,1993. 3. Bajaj, K.K., Oil-mist lubrication of high temperature paper machine bearings, Lub. Eng., 50, 564568, 1994. 4. Alemite Oil Mist Application Manual, Alemite Corporation, Form 37-88, 3-31, 1992.

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39 CIRCULATING OIL SYSTEMS E. R. Booser CONTENTS Reservoirs..............................................................................................................................................404 System Piping........................................................................................................................................407 Pumps.....................................................................................................................................................408 Sizing Example.....................................................................................................................................409 Coolers...................................................................................................................................................410 Filtration and Purification...................................................................................................................410 Instrumentation and Control.............................................................................................................411 References.............................................................................................................................................412

A circulating oil system is generally brought into use where equipment needs, cooling, or reliability requirements preclude simpler lubrication supplies. These circulating systems usually involve an oil reservoir, piping, pumping, cooling, and filter elements. While Figure 1 shows a typical arrangement, there are wide variations. The reservoir, for instance, might simply be the oil sump for a gasoline engine or an industrial gas turbine. Characteristics of typical systems and their components are listed in Table 1 and discussed in this chapter.1-3

RESERVOIRS The reservoir is the core of the system in serving as a storage vessel, settling tank, foam separator, and deaerating chamber. The reservoir top surface is frequently used for mounting pumps, coolers, filters, controls, and instrumentation. See Figure 2 for a typical layout. RESERVOIR CAPACITY The system lubricant volume is the total feed requirement of all bearings, gears, controls, and other machine elements at maximum flow conditions multiplied by the oil dwell time. For a paper mill dryer section requiring 20 gpm feed and a 40-min dwell time for settling of contaminants (see Table 1), for instance, the working capacity becomes 800 gallons. Variables such as thermal expansion, foam, and air venting require extra free space above the normal oil level to assure that the reservoir does not overflow and oil return lines are not blocked. This allowance normally comprises about 1 min of flow rate or a minimum of 4 to 8 in. of height. PROPORTIONS Low oil depth permits faster escape of entrained air and quicker settling of water and solids. A long tank desirably enables locating the pump suction farther from the returning oil. The most Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 Typical system arrangement. (From Twidale, A. J. and Williams, D. C. J., in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 395.) suitable proportions are commonly width = height = length/2. Available headroom and floor area also frequently influence reservoir proportions. CONNECTIONS Oil draining back from bearings and gears should return at one end of the reservoir and just above the oil level to minimize foam and to release entrained air. Any line returning air-free oil from a pump pressure-relief valve or oil conditioning system should extend 6 in. or more below the reservoir oil level to avoid introducing air into the bulk oil. Internal reservoir baffling, by lengthening the path to the oil pump and minimizing stagnant areas, then promotes separation of entrained air and contaminants. Sloping the reservoir bottom from 1 in 15 to 1 in 30 diverts water and impurities to a drain at the low end for removal to the purification system. Oil-pump suction must be kept well below the lowest oil level to avoid sucking air and losing pump prime, but should be 6 in. or more above the Copyright © 1997 CRC Press, LLC.

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FIGURE 2 Typical reservoir layout. (From Twidale, A. J. and Williams, D. C. J., in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 395.) reservoir bottom to avoid pickup of dirt and water. Where oil level varies widely, a floating suction with attached strainer will serve. RESERVOIR HEATING Either electric or steam immersed heaters should be sized to heat the oil from ambient temperature to that required for starting, often in the 70 to 100°F range (20 to 40°C), within 4 hours. For steam heating coils, the pressure should not exceed 50 psig (equivalent to 300°F) to avoid oil decomposition on the heater surface. Surface temperature with electric heaters is controlled by limiting heat flux to about 8 W/in.2 for low viscosity oils in the VG 32- to VG 68-range and 4 watts above VG 250.1 Copyright © 1997 CRC Press, LLC.

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SYSTEM PIPING Pipes must be large enough to prevent cavitation in pump suction lines, to avoid undue pressure drop in feed lines, and to avoid back-up in drain lines at the minimum oil operating temperature. As indicated in Figure 3, feed lines usually operate at 5 to 10 ft/s This may sometimes be raised to 20 ft/s for low-viscosity hydraulic oils and to keep contaminants from separating in transit. For high-viscosity oils in long lines, velocities below 5 ft/s are desirable to minimize the pressure drop.2

FIGURE 3 Chart for approximating feed and drain line sizes. (From Booser, E. R. and Smeaton, D. A., in Standard Handbook of Lubrication Engineering, McGraw-Hill, New York, 1968, 23. With permission.) Feed line flow is characterized by the dimensionless Reynolds number:

where flow is Q gpm, d is pipe i.d. in inches, and v is centistoke oil viscosity from Figure 4. For Reynolds numbers up to 2000, flow is viscous and pressure drop Pf psi in a feed line L feet long becomes:

where oil density p lb/in.3 is commonly about 0.0307. For turbulent flow when Reynolds number exceeds 2000, pressure drop is given by:

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FIGURE 4 Viscosity at operating temperature for pipe sizing. (From Booser, E. R. and Smeaton, D. A., in Standard Handbook of Lubrication Engineering, McGraw-Hill, New York, 1968, 23. With permission.) Drain lines should not run more than half full to allow space for foam and escape of air, both entrained in the oil and pulled along by the flowing oil. About 1 ft/s is a common full-drain velocity (2 ft/s for running half full) for VG 32 oils with a drain sloped 1 in 40 toward the reservoir. A greater slope in the 1 in 10 range would be appropriate with the more viscous oils in steel- and paper-mill systems. Minimum drain line slope for running half full is obtained from the following relation, using the viscosity from Figure 4 at the minimum operating temperature:

where s is the drain-line slope. While minimum slope is commonly 1 in 40 and is the basis of Figure 5, all available drop is often employed.

FIGURE 5 Drain line sizing. (From Twidale, A. J. and Williams, D. C. J., in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 395.) Copyright © 1997 CRC Press, LLC.

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PUMPS Gear pumps are widely used for oil supply with industrial equipment, automobile and aircraft engines, turbines, and compressors involving volumes of up to 250 gpm or more at discharge pressures to over 1000 psi. Screw-type positive displacement pumps are available for delivery up to 1000 gpm and 1000 psi pressure; since some can accommodate moderate axial displacement, they are often conveniently coupled in-line to the end of a machine shaft. Both gear and screw pumps are limited to a maximum suction lift of about 24 ft. Centrifugal pumps can deliver much larger volumes at moderate pressures up to 175 psi and will handle dirty oil, but they do require a flooded suction and oil viscosity below 250 cSt. While they can operate with restricted output, protection against overheating under a no-flow condition requires continuous bleed through a restrictor in a bypass line. Feed pressure, P (psi), required from the pump is given by:

where Po is psi pressure in the reservoir; P2 is the required oil pressure to be delivered at bearings or other machine elements, commonly in the 10 to 30 psi range; ∆h is the delivery height in inches above the reservoir level for oil of density p (lb/in.3); and ΣPf is the flow friction pressure drop through piping, coolers, and filters. In many systems a filter pressure drop of about 10 psi and a cooler drop of 5 to 10 psi should be included. Power, H, needed to drive the pump in delivering Q gpm is:

where pump efficiency e is commonly 0.6 to 0.8.

SIZING EXAMPLE As an example of sizing a lubrication system, consider a machine requiring 200 gpm of ISO VG 32 light turbine oil at 115°F and 30 psig for a bearing supply manifold 22 feet above an oil reservoir. For a dwell time of 5 min, reservoir capacity would be set at 1000 gallons with 10% free space above the static oil level. Figure 3 indicates that a feed pipe i.d. of 3 in. is adequate. Since VG 32 oil at 115°F has a viscosity of 26 cSt, Reynolds number Re to characterize the flow is given by Equation 1 as:

Since Re is greater than 2000, flow is turbulent. Pipe length is 40 ft. Equivalent length of three elbows and a tee fitting is given from Table 2 as 3(25) + 60 = 135 diameters, or 34 ft. With a total equivalent length of 40 + 34 = 74 ft, Equation 3 gives a turbulent pressure drop of

From Equation 5, the total pressure P to be supplied by the pump with allowance of 10 psi pressure drop through the cooler and 10 psi through the filter is:

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Assuming pump efficiency of 0.7, power needed to drive the pump from Equation 6 is:

To accommodate oil draining at 70°F during starting when viscosity of VG 32 oil is 75 cSt, the drain line i.d. required from Figure 5 is 5 in. for running half full with a slope of 1 in 40.

COOLERS Friction loss from bearings, gears, pumps, and contact with high temperature surfaces all combine to heat oil returning to the reservoir, commonly to the 50 to 70°C (120 to 160°F) range with conventional petroleum circulating oils; higher temperatures lead to shortened oil life. An oil cooler positioned between the main oil pump and the lubricated elements then uses water or air to drop the oil by 10 to 20°C (20 to 40°F) for feeding to bearings and other machine components at 45 to 60°C (105 to 140°F). Various plate, fin, and tube coolers are used in small oil-circulating systems, while shell-and-tube coolers are usually employed in larger systems. Oil pressure drop through the shell is commonly 6 to 15 psi, while cooling water pressure drop through the tubes is 2 to 5 psi. When water is unavailable, air is commonly blown by a fan over tubes carrying the hot oil in a radiator.

FILTRATION AND PURIFICATION Degree of oil cleanliness influences the type of oil purification indicated in Table 1. Filtration down to 1 micron (1 millionth of a meter [1 mm] or 0.00004 in.) may be needed with some hydraulic valves and actuators, while 150-micron strainers suffice in steel rolling mills with their rugged construction and frequent maintenance. Filters, the most common means for cleaning additive-type oils, are generally disposable cartridges of pleated paper, cotton bags, or waste cellulose or fine mesh screens. Activated clay or alumina cartridges are used to adsorb small amounts of water and oxidation products with phosphate ester fireresistant fluids and some uninhibited petroleum oils, but these active filters should be avoided with additive-type oils since they tend to remove rust and oxidation inhibitors. Up to 1% water contamination can be removed with blotter type cartridges; extensive water removal requires a separate purification unit. Frequently a combination of techniques are advantageous. Settled water and sludge collecting on the sloping reservoir bottom can be removed in a batch process. A centrifuge in a by-pass circuit processing 10% of the oil per hour effectively removes water and dirt particles down to l-µm size. Magnets can be incorporated with dual basket filters using a changeover valve. Mechanically cleaning filters employ interleaved radial metal plates which plough the dirt from gaps between metal discs when the filter pack is rotated. Copyright © 1997 CRC Press, LLC.

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INSTRUMENTATION AND CONTROL Table 3 and the following considerations reflect controls and alarms needed to ensure continuous feed of sufficient lubricant to bearings, gears, and related machine elements.

RESERVOIR INSTRUMENTATION The simplest device to indicate proper oil level is a sight gauge. Large reservoirs may use high-level and low-level mercury float switches to sound an alarm. A cooler water leak would trip the high-level switch; an oil leak or excessive oil consumption would sound the low-pressure alarm. A thermocouple or resistance temperature detector would indicate temperature at a central control console. A thermally actuated switch with electric immersion heaters or a temperature-actuated valve with steam heating would be used to bring the oil to a suitable start-up temperature. When oil viscosity reaches operating condition, closing a second temperature-actuated switch makes the plant operative. PUMP CONTROLS A pressure-relief valve at their output protects positive displacement pumps from overload. No such protection is ordinarily needed with centrifugal pumps which accommodate broad changes in flow demand with little variation in delivery pressure. If a lubricated machine must continue to run after failure of the main pump, one or even two back-up oil feed supplies must be automatically started by flow- or pressure-operated switches. As an alternative, as in steel mill systems, an elevated emergency supply or pressure tank downstream of the cooler can provide a diminishing oil supply for up to 5 min (see Figure 6). Sized to accommodate emergency oil quantity Q with the tank 2/3 full, the upper air cushion V1 maintains the pressure at that of the pump discharge. In case of a pumping system failure, this air cushion drives enough oil for a satisfactory shut-down (as in clearing of hot steel in a steel mill). Since some air is absorbed in the oil, air must be added from time to time. To avoid accidental over-pressurization, an air regulator is installed along with a safety valve set to open 10 psi above the pump relief-valve setting.

COOLER DEVICES To assure oil delivery at the proper temperature, a temperature sensor in the cooler outlet oil flow can be used to control pneumatically or electrically an inlet water valve to the cooler. INSTRUMENTS AT LUBRICATED MACHINERY Pressure and temperature should be monitored in bearing headers with switches to give low-pressure and high-temperature alarms. In critical systems, automatic controls back up these alarms and start

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FIGURE 6 Pressure tank for oil supply during emergency shutdown. (From Twidale, A. J. and Williams, D.C. J., in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 395.) spare or emergency oil pumps. Secondary switches will trip out the lubricated equipment if temperature rises or pressure falls to the danger zone. Resistance temperature detectors or thermocouples at each bearing and in oil drains from bearings and gears provide a useful check on machine operation. When a monitored bearing operates 10 to 15°C (20 to 25°F) hotter than normal, warning signals or even automatic shutdown should be provided.

REFERENCES 1. Twidale, A. J. and Williams, D.C. J., Circulating oil systems, in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 395–409. 2. Booser, E. R. and Smeaton, D. A., Circulating-oil-system design, in Standard Handbook of Lubrication Engineering, McGraw-Hill, New York, 25, 1968, 23–44. 3. Wilcock, D. F. and Booser, E. R., Lubrication techniques for journal bearings, Machine Design, Vol. 59 (15), 1987, 84–89.

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V Friction, Wear, and Surface Characterization

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40 Surface Texture V. V. Dunaevsky, Y.-R. Jeng, and J. A. Rudzitis CONTENTS Summary and Tips on Surface Texture......................................................................................415 Surface Texture Characterization................................................................................................417 Measurement of Surface Texture................................................................................................422 Contact Interaction of Surfaces..................................................................................................423 Effects of Surface Texture on Machine Components............................................................430 Acknowledgment...........................................................................................................................433 References.......................................................................................................................................433

SUMMARY AND TIPS ON SURFACE TEXTURE METROLOGICAL ASPECTS All engineering surfaces are not perfect. Due to manufacturing causes and physics of solids, they have macro- and microgeometrical deviations from the nominal dimensions which are called surface texture and surface topography. Figure 1 illustrates the relative size of surfacerelated phenomena.

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These deviations have varied magnitude and patterns, and they affect various functional characteristics of machine components including wear, life, and reliability. For convenience of analysis, a measured or total profile of the surface is attenuated through mechanical or electronic filtering into several individual waveforms called waviness and roughness, of which roughness has a smaller wavelength. The individual waveforms are represented through a set of geometrical characteristics which are known as amplitude, spacing, and hybrid parameters of surface texture. Some standards for surface texture are given in Table 1.

FUNCTIONAL ASPECTS Because of surface texture and regardless of its magnitude, except for very smooth surfaces, actual contact area between solids is very small and in all situations does not depend on nominal area.

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Because of smallness of the real area of contact, friction temperature and contact pressure are very high. In general, the smaller the roughness height, the less is the thickness of lubricant film necessary to support the load. Average roughness height parameter, Ra, which is also known as center line arithmetic average (AA) and center line average (CLA), is the most popular surface texture assessment characteristic which in many occasions interrelates with other roughness parameters. Typical values of roughness height for a large variety of mating surfaces of precision, sliding and sealing members are in a range of 0.025 to .6 µm Ra with the lowest values related to bearing balls and gauges. The original magnitude and pattern of the surface texture of a component tend to change in a process of contact interaction with other components by adapting to the operational conditions. Lay is an important surface texture characteristic.

SURFACE TEXTURE CHARACTERIZATION Schematics of departure of roughness profile from an ideal shape: a. Theoretical profile of turned surface. b. Random component due to physical processes developed on the interface between tip of the tool and work-piece. c. Resulting profile.

The randomness of a surface profile dictates a statistical approach in the analysis of surface finish. Hence, all surface parameters are statistical parameters.

FIGURE 2 Pictorial view of engineering surface.

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FIGURE 3 Distortion of the measured profile in its graphical representation due to nonequal horizontal and vertical magnifications.

FIGURE 4 Statistical nature of the surface finish. PARAMETERS DERIVED FROM ABBOTT-FIRESTONE CURVE, AFC (BAC)15,16 These parameters, which are calculated using Rk filtering, are referred to a datum line computed from the AFC and intersecting with the ordinates at bearing ratio zero and 100%. This line separates the surface profile into three parts:

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• Material-filled profile peak area (A,) — a measure of the amount of material that will be removed in the run-in period. • Core roughness area of profile. • Profile depth or profile valley area (A2) — a measure of the area in the profile that can retain lubricant.

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Within these areas the following parameters are distinguished: • Rk — core roughness depth. The depth of the roughness profile excluding prominent peaks and grooves. It will, after initial run-in period, carry the load and influence life and performance. • Rpk — reduced peak height. The height of the top portion of the surface profile, i.e., of the profile peaks exceeding the core profile. It is often worn away during the run- in period. • Rvk — reduced valley depth. The depth of the lowest part of the surface profile, i.e., of grooves extending below the core profile. It is a measure of the oil-retaining capacity of the surface in question. • Mr1 — material component relative to peaks. The material ratio (tp) at which Rpk and Rk meet. This is the upper limit of the core roughness profile. • Mr2 — material component relative to valleys. The material ratio (tp) at which Rvk and Rk meet. This is the lower limit of the core roughness profile.

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SURFACE TEXTURE SYMBOLSA (PER ANSI Y 14.36 — 1978)

MEASUREMENT OF SURFACE TEXTURE Cutoff length or sample (sampling) length or roughness width cutoff (RWC), which is usually = 0.8 mm (0.03 inch) define the filter. The other standard cutoff values are 0.08 mm (0.003 inch), 0.25 mm (0.01 inch), 2.5 mm (0.1 inch), 8 mm (0.3 inch), or 25 mm (~1 inch). The upper scale profilometers allow also customary values of cutoff. Waves shorter than cutoff length are classed as roughness, R. Waves longer than the cutoff length are classed as waviness, W. All roughness parameters are measured within a certain number of sampling lengths — typically five. Waviness and total parameters are also measured on a specified traversing length. A magnitude of the RWC does not depend on a type of filter among which a so-called Gaussian filter, which is both ISO and ANSI/ASME recommended, distinguishes waviness from roughness more sharply than other filters such as, e.g., analog 2RC and digital phase-correct filter 2RC. Selection of the proper cutoff length should be based upon surface being measured.

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a. For periodic surfaces, e.g., turned surface, the value of cutoff is based on the spacing feature parameter, Sm, mean spacing, as defined by Table 7.7 b. For random, nonperiodic type surfaces, e.g., ground surface, select the proper cutoff based upon Ra, roughness average, per Table 8.7

CONTACT INTERACTION OF SURFACES CONTACT OF IDEAL AND ROUGH SURFACES10

A diagram of a contact interaction of the ideal and rough surfaces: 1, plastically deformed asperities; 2, elastically deformed asperities. P, load; aCR, deformation of the asperities at which plastic flow of the asperity commences. I, initial position of the ideal rigid flat; II, position to which a rigid flat was transferred, producing elastic and plastic deformations of the asperities; S, distance between II and mean line m (separation between the surfaces). The load is supported by those asperities (shaded) whose heights are greater than the separation u between the planes. Copyright © 1997 CRC Press, LLC.

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FIGURE 5 Roughness ranges of production processes (NSI B46.1—1985).

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CONTACT OF TWO ROUGH SURFACES19

(a) Geometric presentation of the contact interaction of two rough surfaces; (b) equivalent surface; L, separation between the surfaces, h, roughness heights of the surfaces; o, mean line. The equivalent surface, ES, has been introduced to simplify a contact analysis of two rough surfaces. ES is such a surface which, after being compressed by a smooth rigid plane, maintains contact area and separation (from the rigid plane) equal to those of two rough surfaces under the same load. See Tables 9 and 10.

In the above: v, n(0)i = number of the intersections of the profile with mean line per unit length for ES and the individual isotropic surface respectively; µ, mi, = number of the profile peaks for ES and an individual isotropic surface, respectively; i = 1,2, indices of the individual isotropic surfaces.

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CONTACT OF IDEAL SURFACE AND ASPERITY A contact of the individual asperity with the ideal surface is often presented as a contact of the hemisphere with a plane: (a) plastic contact; (b) elastic contact.

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Surface Texture

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EFFECTS OF SURFACE TEXTURE ON MACHINE COMPONENTS

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ACKNOWLEDGMENT The author gratefully acknowledges the contribution to this work, as it relates to the sections of surface roughness measurement and filtering, stemming from technical discussions with Ira Kerns, Precision Devices, Inc., and Richard Wismer from Measuretech, representing Hommel Werke.

REFERENCES 1. Williamson, J.B.P., The shape of the surfaces, CRC Handbook of Lubrication, Theory and Practice of Tribology, Vol. 2, Theory and Design, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1984, 3–16.

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2. Surface Texture (Surface Roughness, Waviness and Lay), ANSI/ASME B46.1 — 1985, American Society of Mechanical Engineering (ASME), New York, 1985. 3. Chung, Y.-W. and Sriram, T.S., Scanning tunneling microscopy, ASM Handbook, Vol. 18, Friction, Lubrication and Wear Technology, ASM International, 1992, 393–398. 4. Garbini, J.L. et al., Surface profile measurement during turning using fringe-field capacitive profilometry, ASME Preprint 90 - WA - DSC - 25, 1990. 5. Moore, D., Principles and Applications of Tribology, Pergamon Press, New York, 1975. 6. Hailing, J., Principles of Tribology, Macmillan Press Ltd., London, 1975, 38–47. 7. Nordberg, M.E., Surface Finish Concepts, Mathematics and Measurement Techniques, Precision Devices, Milan, MI, 1993. 8. Dunaevsky, V.V., Measurement of local microscopic wear, J. Tribol., 108, 35–41, 1986. 9. Peters, J. et al., “Assessment of Surface Typology Analysis Techniques,” Annals of CIRP 28/2, 1979. 10. Rudzitis, J. A., Microgeometry and Contact Interaction of Surfaces, Zinatne, Riga, Latvia, 1975,12–38. 11. Kragelsky, I.V. and Michin, N.M., Handbook of Friction Units of Machines, Dunaevsky, V.V., Ed., ASME Press, New York, 1988, 8–11. 12. Dunaevsky, V.V., Research on the Relationship between the Type and Quality of Production Processes and Wear of the Main Components of Cylinder/Piston and Valve Groups of the Internal Combustion Engines, Doctoral thesis, Riga Technical University, Riga, Latvia, 1975, 72–82. 13. Greenwood, J.A. and Williamson, J.B.P., Contact of nominally flat surfaces, Proc. R. Soc. London, Ser. A, 295, 300–319, 1966. 14. Kragelsky, I.V., Friction and Wear, Pergamon Press, London, 1965, 25. 15. Drews, W. and Weniger, W., Rediscovering the Abbott-Firestone curves, Quality, pp. 50–53, Sept. 1989. 16. Measurement of Surface Roughness, Parameters Rk, Rpk, Rvk, Mr1, Mr2 for the Description of the Material Portion (Profile Bearing Length Ratio) in the Roughness Profile, DIN 4776—1985, Beuth Verlag GmbH, Berlin, 1985. 17. Sander, M., A Practical Guide to the Assessment of Surface Texture, Feinpruf Perthern, Göttingen, Germany, 1989, 35–55. 18. Marks’ Standard Handbook for Mechanical Engineers, 8th ed., McGraw-Hill, New York, 1979, 13.73–13.79. 19. Dunaevsky, V.V. and Rudzitis, J.A., Determination of geometrical parameters of an equivalent surface when studying the process of contact interaction of rough surfaces, Machinery Construction, No. 3, Mashinostroenie, Moscow, 1975, 66–70. 20. Bowden, F.P. and Tabor, D., Friction and Lubrication of Solids, Clarendon Press, Oxford, 1986, 20. 21. Moore, D.F., The Friction and Lubrication of Elastomers, Pergamon Press, New York, 1972, 224–227. 22. Piggot, M.R. and Wilman, H., Nature of the wear and friction of mild steel on mild steel and the effect of surface oxide and sulfide layers, Conference on Lubrication and Wear, Institution of Mechanical Engineers, London, 1957, 613. 23. Jeng, Y.-R., Experimental study of the effects of surface roughness on friction, Tribol. Trans., 33, 402–410, 1990. 24. Shneyder, Yu G., Development of the Regular Microrelief on the Machine Members and Its Operational Characteristics, Mashinostroenie Publishing House, Leningrad, 1972, 21. 25. Lubricomp Internally Lubricated Reinforced Thermoplastics and Fluoropolymer Composites, Bull.254, LNP Engineering Plastics, Exton, PA, 1994, p. 4. 26. Husu, A.P., Vittenberg Yu R., and Pal’mov, V.A., Surface Roughness Theory Probability Approach, Nauka, Leningrad, 1975, 15. 27. Reshetov, D.N., Machine Design, Mir Publishers, Moscow, 1978, 104–115. 28. Tallian, T.E., Rolling contact failure control through lubrication, Proc. Inst. Mech. Eng., 182, 205–236, 1967–1968. 29. Dunaevsky, V.V., Surface Texture and Topography. Measurement, Characteristics & Application, Allied Signal/Truck Brake System Co., Elyria, OH, 1995. 30. Mummery, L., Surface Texture Analysis. The Handbook, Hommelwerke, New Britain, CT, 1991.

FURTHER READING An extensive list of references on the subject of surface texture and topography can be found in Reference 3 of the ASM Handbook and in a section on Surface Topography in Mechanical Engineer’s Reference Book, 12th ed., Smith, E.H., Ed., SAE, Warrendale, PA, 1994, 125–132.

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41 Typical Friction and Wear Data A. W. Ruff CONTENTS Metal and Polymer Friction and Wear Data..............................................................................435 Ceramic Friction and Wear Data.................................................................................................435 Abrasive Wear Standard Test Data.............................................................................................435

METAL AND POLYMER FRICTION AND WEAR DATA This set of data was extracted from the ACTIS Tribomaterials I Database, NIST Reference Database 22 (1989). The entire database includes about 370 records covering metals, polymers, composites, and ceramics. The data sources were archival publications, and roundrobin activities. For more complete information, NIST Database 22 or the original sources should be consulted.

CERAMIC FRICTION AND WEAR DATA This set of data was extracted from the ACTIS Tribo-ceramic Materials Database, NIST Reference Database 47 (1994). The entire database includes about 370 records, covering 7 different ceramic materials in 36 combinations. The data sources were NIST research projects, archival publications, and round-robin activities. For more complete information, NIST Database 47 or the original sources should be consulted.

ABRASIVE WEAR STANDARD TEST DATA: ASTM G-65 DRY SAND/RUBBER WHEEL TEST These summary data were obtained from a series of interlaboratory tests done by members and associates in ASTM Committee G-2. More details can be found in the ASTM Standard Test G-65 and the associated Research Report G2-1004, both of which are available from ASTM, Philadelphia, PA.

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42 Friction and Wear Equations Valery V. Dunaevsky CONTENTS Friction Equations..........................................................................................................................445 Influence of the Various Factors on Coefficient of Friction (Elastoplastic Contact)..........448 Wear Equations...............................................................................................................................449 Example of Use of Wear Equations...........................................................................................450 Hardness Considerations..............................................................................................................452 References.......................................................................................................................................453

FRICTION EQUATIONS

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INFLUENCE OF THE VARIOUS FACTORS ON COEFFICIENT OF FRICTION7 (ELASTOPLASTIC CONTACT)

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WEAR EQUATIONS

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EXAMPLE OF USE OF WEAR EQUATIONS A journal bearing consists of a horizontal stainless steel shaft (Figure 3), of a diameter 25 mm, rotating in a copper bushing with the initial diametral clearance of 0.05 mm. The shaft rotates with a speed of 500 rpm under a normal load of 10 N (1.02 kg). Penetration hardness of the bushing is 50 HBN (Brinell hardness number), i. e., 50 kg/mm2. Copyright © 1997 CRC Press, LLC.

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It is required to select a regime of lubrication at which the bearing can operate no less than 5000 h before reaching a clearance of 0.5 mm, at which the bearing is considered inoperable.

FIGURE 3 A journal bearing consisting of a rotating shaft and stationary bushing. Solution: 1. Adhesive wear equation is used:

2. Considering that only wear of the bushing contributes to the increase of the bearing clearance, a critical worn volume of the bushing is

where t = 0.5 ⋅ (cf - ci) — thickness of the worn layer; cf = final clearance; ci = initial clearance; l = length of a bushing; dav = average diameter of worn layer.

3. Length of the sliding distance

where ω = shaft velocity, rpm; n = duration of the operation of the bearing, min. 4. Substituting ( 2 ) and ( 3) into Equation 1, obtain

5. Find k from (4) which provides 5000 h of operation

6. Substituting into (5) the related values of the parameters in the compatible units, obtain:

7. Comparing the above value of k with k values in a section on adhesive wear Ref. 38, a column for metals with intermediate metallurgical compatibility to which shaft and bushing materials belong,39 find that the k from Equation 6 is close to the value of k for good lubrication, i.e., 2 ⋅ 10-6.

It is recommended to use good lubrication to provide 5000 h of the required bearing life.

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Consequently, as a word of caution in using the demonstrated technique for lubrication and wear analysis, wear coefficients may vary significantly, up to several magnitudes, depending on the test conditions and materials compatibility. Besides, the wear equation we used may not necessarily be adequate to the actual operating conditions. This demonstrated technique should be only an approximate and preliminary tool for wear analysis which (in a demanding application) should be verified through bench testing or testing of the actual components under conditions approaching or representing the real operation. HARDNESS CONSIDERATIONS In dealing with friction and wear problems, it is often necessary to use indentation (penetration) hardness (i.e., the ratio of load applied to area of indentation produced by plastic yielding). The indentation hardness is best measured by a Vickers, Knoop, or Brinell test. Brinell hardness numbers are equal to hardness stress expressed in kg/mm2. The indentation hardness is about three times the yield stress in uniaxial tension or compression. The Rockwell hardness test is a test of the indentation type, but it measures the vertical distance through which the indenter (generally a cone with rounded end) moves during indentation. Rockwell hardness number R relates to a penetration hardness H by means of the Equations:36

where k1 and k2 are constants. The conversion between Rockwell and penetration hardness is shown graphically, Figure 4. Note that there are several Rockwell scales, depending on load and indenter size and shape used.

FIGURE 4 Diagram to convert Rockwell hardness numbers into hardness stress, H, N/mm2. (A) A scale, diamond brale indenter, 60-kg load; (B) B scale, 1/16” steel ball, 100-kg load; (C) C scale, brale indenter, 150-kg load; (D) D scale, diamond brale indenter, 100 kg load; 15-N, N scale for a superficial test using a diamond brale with a 15-kg load. Copyright © 1997 CRC Press, LLC.

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REFERENCES 1. Leonardo Da Vinci, About Myself and My Science (1508). Selective works, 1, Academia, Moscow, 1932. 2. Kraghelsky, I.V. and Schedrov, V.S., Evolution of the Science of Friction [in Russian], AN USSR (USSR Academy of Sciences), Moscow, 1956, 1–69. 3. Amontons, G., On the resistance originating in machines, [in French], Mem. Acad. Roy., 1699, 206–222. 4. Parent, A., New static with and without consideration of friction, or rules for determination of friction forces in machines [in French], Mem. Mathem. Phys. Acad. Sci., Paris, 1704, 173–197. 5. Rabinowicz, E., Friction and Wear of Materials, John Wiley & Sons, New York, 1965, 66. 6. Coulomb, C.A., The theory of simple machines [in French], Mem. Math. Phys., 10, Paris 1785, 161–331 7. Kraghelsky, I.V. and Vinogradova, I.E., Coefficients of Friction [in Russian], Mashgiz, Moscow, 1962, 5–26 8. Ernst, H., and Merchant, I.E., Surface friction between metal — a basic factor in metal cutting process, Proc. Spec. Summer Conf. Friction and Surface Finish, M.I.T., Cambridge, MA, 1940, 76–101 9. Merchant, M.E., The mechanism of static friction, J. Appl. Phys., 11, No. 3, 1940. 10. Kraghelsky, I.V,. Friction and W ear, Pergamon Press, London, 1965. 11. Kraghelsky, I.V., Evaluation of the friction properties of materials in sliding contact, Industrial Lab., 34(8), 1007–1009, 1968. 12. Moore, D.F., Principles and Applications of Tribology, Pergamon Press, New York, 1975 13. Ettles, C.M., The thermal control of friction at high speeds, A SME Trans., J. Tribol., 108, 98–107, 1986 14 Dine, O.C. and Ettles, C.M., et al., Some parameters affecting tactile friction, ASME Trans., J. Tribol., 113, 512–517, 1991 15. Euler, L., Friction of Solids [in French], Hist. Acad. Roy., Berlin, 1748. 16. Petroff, N., Friction in machines and the effect of the lubricants, (a) In Russian, Eng. J., St. Petersburg, pp. 71–140, 228–274, 377–436, 1883; (b) German translation by L. Wurzel, L.Voss, Hamburg, 1887, pp. 187 17. Fuller, D.D., Theory and Practice of Lubrication for Engineers, John Wiley & Sons, New York, 1956, 12–26. 18. Sneck, H.J. and Vohr, J.H., Hydrodynamic lubrication, CRC Handbook of Lubrication (Theory and Practice of Tribology), Vol. 2, Theory and Design, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1984, 72, 73, 90. 19. Harris, T.A., Friction and wear of rolling element bearings, ASM Handbook, Vol. 18, Friction, Lubrication, and Wear Technology, ASM International, Metals Park, OH, 1992, 499–514. 20. Errichello, R., Friction, lubrication, and wear of gears, ASM Handbook, Vol. 18, Friction, Lubrication, and Wear Technology, ASM International, Metals Park, OH, 1992, 536–545. 21. Archard, J.F., Contact and rubbing of flat surfaces, J. Appl. Phys., 24, 1953, 981–988. 22. Hailing, J., Toward a mechanical wear equation, ASME Trans., J. Tribol, 105, 212–220, 1983 23. Rabinowicz, E., Abrasive wear resistance as a material test, Lubr. Eng., 33 (7), 378–381, 1975 24. Evans, A.G. and Marshal, D.B., Wear mechanisms in ceramics, Fundamentals of Friction and Wear of Materials, Rigney, D.A., Ed., ASM, Metals Park, OH, 1980, 439. 25. Kruschov, M.M. and Babichev, M.A., Method of wear testing of metals by friction against abrasive surface [in Russian], Friction and Wear in Machines, 1, USSR Academy of Sciences, Moscow, 1941. 26. Kruschov, M.M., Resistance of metals to wear by abrasion related to hardness, Inst. Mech. Eng., Conf. Lubrication and Wear, London, 1957, 655–659. 27. Kraghelsky, I.V, et al., Fundamental of friction and wear analysis [in Russian], Mashinostroenie, Moscow, 1977, 321–322. 28. Kruschov, M.M., Principles of abrasive wear, Wear, 28, 69–88, 1974 29. Kruschov, M.M. and Babichev, M.A., Investigation into the Wear of Metals [in Russian], USSR Academy of Sciences, Moscow, 1960. 30. Kraghelsky, I.V. and Michin, N.M., Friction Units of Machines. Handbook, Dunaevsky, V.V., Ed., ASME Press, New York, 1988, 11–16.

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31. Dunaevsky, V.V., Measurement of local microscopic wear, ASME Trans., J. Tribol., 108(1), 35-41, 1986. 32. Shneyder, E. W. et al., Effect of Speed and Power Output on Piston Ring Wear in a Diesel Engine, No. 880672, SAE Tech. Pap. Ser., Warrendale, PA. 33. Chen, G. et al., Wear prediction for unlubricated piston rings, Wear of Materials, 2, 645-647,1989. 34. Stiffer, A.K., Melt friction and pin-on-disc devices, ASME Trans., J. Tribol., 108(1), 105-108, 1986. 35. Bayer, R. G. et al., Handbook of Analytical Design for Wear, Plenum Press, New York, 1964, 67. 36. Rabinowicz, E., Wear coefficients — metals, Wear Control Handbook, ASME Press, New York,1980, 479. 37. Dunaevsky, V.V., Surface topography and surface texture, Tribology Data Handbook, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1997, chap. 40. 38. Dunaevsky, V.V., Generalized wear coefficients, Tribology Data Handbook, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1996, chap. 43. 39. Rabinowicz, E., Wear coefficients, CRC H andbook of L ubrication, Vol. 2, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1984, 204. 40. Mulhearn, T.O. and Samuels, L.E., The abrasion of metals: a model of the process, Wear, 5, 478498, 1969.

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Examples of application of wear coefficients k in engineering analysis are shown in References 3 and 22. A specific linear wear rate ih, Equation 11 in a section of wear equations,22 is another example of the dimensionless wear constants. Along with generalized wear coefficients, popular measures of wear rates include wear factors (rates) taken as

where h = thickness of worn layer, m; P = average operating pressure, lb/in.2 (1 lb = 4.45 N; 1 in. = 0.0254 m); V = velocity, ft/min (1 ft = 0.304 m); T = operating time, h. The actual values of k3for various thermoplastic and fluoropolymer composites are given in References 25 and 26. For many applications a wear rate exceeding 2.5 ⋅ 10-7 m/h is considered excessive. One should note that wear coefficients (wear rates) are highly dependent on test conditions (specimen configuration, temperature, environment, and other factors). Consequently “...large errors may arise if the end use conditions differ significantly from those under which data were taken.”17

REFERENCES 1. Peterson, M. B., Design considerations for effective wear control, Wear Control Handbook, ASME, New York, 1980, 443. 2. Archard, J. F, Contact and rubbing of flat surfaces, J. Appl. Phys., 24, 981–988, 1953. 3. Rabinowicz, E., Wear coefficients, CRC Handbook of Lubrication, Vol. 2, CRC Press, Boca Raton, FL, 1984, 201–208. 4. Rabinowicz, E., Friction and Wear of Materials, John Wiley & Sons, New York, 1995. 5. Archard, J. F. and Hirst, W, The wear of metals under unlubricated conditions, Proc. R. Soc. London, Ser. A, 236, 397–410, 1956. 6. Spurr, R. T. and Newcomb, T. P., The friction and wear of various materials sliding against unlubricated surfaces of different types and degrees of roughness, Proc. Conf. Lubrication and Wear, Institution of Mechanical Engineers, London, 1957, 269–275. 7. Avient, B. W. E., Goddard, J., and Wilman, M., An experimental study of friction and wear during abrasion of metals, Proc. R. Soc. London, Ser. A, 258, 159–180, 1960. 8. Lopa, M., A Study of the Influence of Hardness, Rubbing Speed and Load on Abrasive Wear, B.S. thesis, MIT, Cambridge, MA, 1956. 9. Kruschov, M. M. and Babichev, M. A., Resistance to abrasive wear of structurally inhomogeneous materials, Friction and Wear in Machinery, 12, ASME, New York, 1958, 5–23. 10. Samuels, L. E., The nature of mechanically polished surfaces: The surface deformation produced by the abrasion and polishing of 70:30 brass, J. Inst. Met., 85, 51–62, 1956.

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11. Toporov, G. V., The influence of structure on the abrasive wear of cast iron, Friction and Wear in Machinery, Vol. 12, ASME, New York, 1958, 39–59. 12. Rabinowicz, E., Dunn, L. A., and Russel, P. G., The abrasive wear resistance of some bearing steels, Lubr. Eng., 17, 587–593, 1961. 13. Rabinowicz, E., Dunn, L. A., and Russel, P. G., A study of abrasive wear three body conditions, Wear, 4, 345–355, 1961. 14. Fein, R., AWN — A proposed quantitative measure of wear protection, Lubr. Eng., 31, 581, 1975. 15. Rowe, C. N., Lubricated wear, in Wear Control Handbook, Peterson, M. B. and Winer, W. O., Eds., ASME, New York, 1980, chap. 6. 16. Rowe, C. N., Lubricated wear, in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 217. 17. Jahanmir, S. and Fischer, T. E., Friction and wear of ceramics, CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 103–118. 18. Rabinowicz, E., Wear coefficients — Metals, Wear Control Handbook, ASME, New York, 1980, 503. 19. Rabinowicz, E., The wear equation of erosion of metals by abrasive particles, Proc. 5th Int. Conf. on Erosion by Solid and Liquid Impact, Cambridge, England, 1979, pp. 38–1 to 38–5. 20. Bhansali, K. J., Wear coefficients of hard-surfacing materials, Wear Control Handbook, ASME, New York, 1980, 380–382. 21. McNab, I. R. and Johnson, J. L., Brush wear, Wear Control Handbook, ASME, New York, 1980, 1091. 22. Dunaevsky, V. V., Friction and wear equations, Tribology Data Handbook, CRC Press, Boca Raton, FL, 1997, chap. 42. 23. D 3702-78, Standard Test Method for Wear Rate of Materials in Self-Lubricated Rubbing Contact Using a Thrust Washer Testing Machine, Annual Book of ASTM Standards, ASTM, Philadelphia, 1990, 46–50. 24. Lewis, R. B., Wear of plastics — evaluation for engineering applications, Pap. 63-WA-325, Winter Annual Meeting ASME, Philadelphia, Nov. 17–22, 1963. 25. LUBRICOMP Internally Lubricated Reinforced Thermoplastics and Fluoropolymer Composites, Bull. 254–691, LNP Engineering Plastics Inc., Exton, PA, 1988. 26. Wolverton, M. P., Friction and wear in plastic components, Mach. Des., September 26, 1991, 82–90.

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44 Friction Temperatures Valery V. Dunaevsky CONTENTS Introduction...................................................................................................................................462 Nomenclature.................................................................................................................................462 Friction Temperatures by Blok...................................................................................................464 Friction Temperatures by Jaeger.................................................................................................464 Friction Temperatures by Archard.............................................................................................468 Friction Temperatures by Rabinowicz.......................................................................................469 Numerical Examples.....................................................................................................................469 Appendixes.....................................................................................................................................471 References.......................................................................................................................................472

INTRODUCTION Shown here are closed form solutions for contact temperature rise due to frictional heating from heat sources both stationary and moving in fresh paths. The referenced formulas provide compatible solutions for identical frictional heating situations. A common speed criterion L is used in all three groups of formulas. True interface temperature is presented as a combination of surface temperatures of both contacting bodies. Tips: • For very small L, results for the moving source tend to be those for a stationary source of the same size and strength. • If L > 2 the difference between the band and square solutions is small. • If the moving body is a not conductor, then a moving source theory is applicable to a stationary body; likewise, if the stationary body is a nonconductor, then stationary source theory is applicable to the moving medium. • If both bodies have finite conductivity, then the slider is heated by the source and cooled by the oncoming cooler portions of the stationary substrate, while the latter is heated both by the source and by conduction from the slider. • Friction temperatures are high at small real contact areas and are localized in thin superficial layers of the surface.

NOMENCLATURE*

* Nomenclature is given in SI units. Other compatible units, e.g., in the CGS systems, are applicable.

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FIGURE 1 Stationary heat supply evenly distributed over a round surface.

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FIGURE 2 Stationary heat supply with semi-elliptical heat distribution over a round surface.

FIGURE 3 Heat source with uniform distribution of heat over a square surface: (a) uniform heat distribution over square; (b) moving source.

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FIGURE 4 Flat surface 2 slides over a protuberance on surface of body 1.

FRICTION TEMPERATURES BY JAEGER2 (Contact arrangement per Figures 5, 6, and 7. Uniform heat sources.)

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FIGURE 6 A long thin square rod 2 slides on the plane 1 and radiates heat from its faces.

FIGURE 7 Band heat source slides on a stationary plane 1.

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FIGURE 8 Moving circular heat source.

FRICTION TEMPERATURES BY RABINOWICZ14 (Circular contact, Figure 8.)

NUMERICAL EXAMPLES NUMERICAL EXAMPLE 1 A numerical example is conducted with employment of the Archard equations and contact arrangement per Figure 8. To apply the results to a practical problem, the proportion of frictional heat supplied to each body must be taken into account. A convenient procedure is to first assume that all the frictional heat liberated at the interface (Q = fNV) is transferred to body 1 and calculate its temperature rise. Then do the same for body 2. The true temperature rise Tav for the two surfaces, taking into account the division of heat between bodies 1 and 2, is given as

Equation 14 is a more critical measure than Equation 13 of an interface temperature rise due to frictional heating. Consider (Reference 7) a circular contact (heat source) 20 mm in diameter (i.e., Copyright © 1997 CRC Press, LLC.

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radius = 0.01 m) formed by a protuberance (body 2 in Figure 8) and flat surface (body 1 in Figure 8). One surface is stationary and another one moves at V = 0.5 m/s. A heat source formed at the interface of both bodies is stationary relative to body 2 and movable relative to body 1. (Either of the surfaces 2 or 1 can be in actual motion.) Both bodies are of plain carbon steel (C = 0.5 %), with one end of the protuberance (which is essentially a cylinder) maintained at 24°C bulk temperature. The coefficient of friction is 0.1, and the load is N = 3000 N. The properties of contacting bodies are (see also Table A2 – 1 in Appendix):

If one can assume that all the frictional energy is conducted into the stationary relative to heat source surface 2, then its temperature is obtained from the stationary source theory. Consequently, using Equation 1 from Table 5,

In turn, if all the frictional energy went into the moving relative to heat source surface 1, then the temperature of this surface can be obtained from the moving heat source theory using Equation 4 from Table 5, since L = 169 >5 is large:

The true temperature rise for the two surfaces is then obtained from either of the Equations 13 or 14.

Subsequently, Tav (15) = 4.4°C and Tav (16) = 4.7°C. To obtain true contact surface temperature, Tcontact’ the bulk temperature, Tb’ must be added to the temperature rise, i.e.,

in the conditions of this example. See also Reference 8, where the total maximum surface temperature rise is obtained by superimposing the local surface temperature rise on the nominal surface temperature rise and background temperature: Ttotal= Tlocal + Tnominal + Tbackground.

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NUMERICAL EXAMPLE 2

NUMERICAL EXAMPLE 3

APPENDIX 1: ADDITIONAL EXPRESSIONS FOR SURFACE TEMPERATURE RISE AT THE FRICTION INTERFACE 1. Hertzian contact.

Considering an elliptical pressure distribution within a round Hertzian contact of sphere and elastic flat surface, and using an analogy of expressions for temperature distribution due to a stationary point heat source on a plane surface, and for elastic normal deflection of a plane surface due to a normal point force acting on the surface, the Francis9 equation for a temperature rise at the center of a Hertzian contact with ellipsoidal power density distribution is 3/8(Q/ K1).

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2. Square source. Intermediate L.

3. Band source.

4. Surface temperature due to multiple heat sources. See Reference 12 for bulk temperature estimation due to a repetitive braking. A bibliographic list on the subject is in Reference 13.

APPENDIX 2

REFERENCES AND RECOMMENDED LITERATURE 1. Blok, H., Theoretical study of temperature rise at surfaces of actual contact under oiliness lubricating conditions, Proc. General Discussion on Lubrication and Lubricants, Vol. 2, Institute of Mechanical Engineers, London, 1937, 222–235. 2. Jaeger, J.C., Moving sources of heat and the temperature of sliding contacts, J. Proc. R. Soc, NSW, 76, 203–224, 1943. 3. Bowden, F.P. and Tabor, D., The Friction and Lubrication of Solids, Clarendon Press, Oxford, 1986, 33–37, (first published, 1950). 4. Archard, J.F., The temperature of rubbing surfaces, Wear, 2, 438–455, 1958–1959.

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5. Greenwood, J.A. and Williamson, J.B.P., Contact of nominally flat surfaces, Proc. R. Soc., Ser. A, 295, 300–319, 1966. 6. Greenwood, J.A., An interpolation formula for flash temperatures, Wear, 150, 153, 1991. 7. Winer, WO. and Cheng, H.S., Film thickness, contact stress and surface temperature, Wear Control Handbook, ASME, New York, 81–141, 1980. 8. Tian, X. and Kennedy, F.E., Contact surface temperature models for finite bodies in dry and boundary lubricated sliding, J. Tribol., 115, 411–418, 1993. 9. Francis, H.A., Interfacial temperature distribution within a sliding Hertzian contact. ASLE Trans., 14, 41–50, 1970. 10. Tian, X. and Kennedy, F.E., Temperature rise at the sliding contact interface for a coated semiinfinite body, J. Tribol, 115, 1–9, 1993. 11. Carslaw, H.S. and Jaeger, J.C., Conduction of Heat in Solids, 2nd ed., Clarendon Press, Oxford, 1959, 269. 12. Dunaevsky, V.V., Prediction of railroad friction braking temperatures: prediction of average bulk and average surface temperatures of railroad wheels and brake discs, Tribol. Trans., 34(3), 343–352,1991. 13. Dunaevsky, V.V., Surface temperature in oscillating rectangular contacts, Preprint 94 - NP - 4G 3, presented at the 49th STLE Annual Meeting in Pittsburgh, May 1–5, 1994, 1–12. 14. Rabinowicz, E., Friction and Wear of Materials, 2nd ed., John Wiley & Sons, New York, 1995, 96–101. 15. Holm, R., Temperature development in a heated contact with application to sliding contacts, J. Appl. Mech., 19, 369–374, 1952.

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LUBRICATION 45 BOUNDARY RELATIONS Richard S. Fein CONTENTS Introduction....................................................................................................................................474 Severity Variables...........................................................................................................................476 Boundary Lubricants.....................................................................................................................478 Boundary Lubrication Determination........................................................................................479 Boundary Lubrication Assessment for Design and Problem Analysis.................................482 References.......................................................................................................................................485

INTRODUCTION Boundary lubrication is the regime of liquid or grease lubrication in which the load is carried on asperities by physical processes which include hydrodynamic flow, elastic and plastic deformation, fracture, adherence, etc. Chemical processes including chemisorption, corrosion, oxidation, etc. interact strongly with the physical processes. Consequently, boundary lubrication depends strongly on the chemical and physical properties of the sliding surfaces, surface texture, chemical and physical properties of the lubricant, and products of chemical reaction of the lubricant with the surfaces and ambient atmosphere.1,2 Table 1 shows the dependence of dimensionless friction and wear coefficients on bearing, lubricant, and atmosphere. The dimensionless wear coefficient is often called the “Archard” or the “Archard-Burwell and Strang-Holm” wear coefficient k.k is the product of the wear volume V and indentation hardness H divided by the product of the load W and sliding distance L; or k is the product of the depth of wear d and indentation hardness H divided by the product of the apparent pressure P and sliding distance L.3 Usually, k is approximately the same for opposing surfaces (within a factor of two) and is often approximately constant over considerable ranges of operating variables and bearing configuration. Table 1 shows over a 50 million-fold range of wear coefficient and less than a 100-fold range of friction coefficient resulting from changing the various combinations of materials under “unlubricated” (i.e., negligible gas-film load support) to boundary-lubricated (i.e., negligible load support by bulk liquid or grease lubricant) sliding. Usually, introduction of nonsimilar but compatible bearing materials,46 oxygen in air, and liquid lubricants reduces both friction and wear. Note a considerable overlap in friction coefficients with those observed in full hydrodynamic (including elastohydrodynamic) lubrication. Figure 1 shows the following three (sub-)regimes commonly observed under boundary lubrication: 1. Mild wear (MW). The mildest combinations of load and sliding velocity produce good wear protection and largely metallo-organic wear products. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 Secondary lubrication regimes. 2. Damage-controlled wear (DCW). An increase in load produces an abrupt transition to higher wear and friction coefficients. DCW wear debris often appears to be largely amorphous metallic oxides, but always includes metallo-organic products. The wear rate, although too high to be tolerated for long, does not very rapidly cause an intolerable bearing geometry change. 3. Surface-damage wear (SDW). A further increase in load produces another abrupt transition to a very rapid surface texture change and/or an intolerable very rapid bearing geometry change. Wear products tend to be metallic and inorganic oxidation products in the case of a large surface texture change, but are likely to be predominantly inorganic reaction products in the case of very rapid wear.

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SEVERITY VARIABLES Load and sliding velocity are just two of a number of variables termed operational severity.1,7,18 An increase in these variables usually increases the likelihood of a transition to a regime of higher wear and friction coefficient. Friction coefficient transitions usually coincide with wear coefficient transitions, but are often too small to observe. Table 2 lists the most common severity variables which, besides promoting transitions, appear to promote modification of sliding surfaces through formation and removal of chemically changed surface films or boundary films. The greater the severity of sliding, the more rapidly boundary films are formed and removed. The consequence is that run-in to a steady-state film thickness and composition occurs more rapidly at higher severity. Figure 2 shows schematically the variation of film formation and removal rates on a previously unslid surface as a function of sliding duration (distance or time) and Figure 3 shows the resulting film thickness.7,18

FIGURE 2 Effect of severity on boundary film formation and removal rates. F, film information rate; R, film removal rate.

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FIGURE 3 Boundary film thickness dependence on sliding distance and severity.

FIGURE 4 Effect of operational severity and chemical reactivity on wear rate.7 Figure 4 schematically illustrates the dependence of wear on material reactivity and operational severity. At low reactivity, insufficient metallo-organic film is formed to adequately shield the substrate-bearing material and the metallic element content of the metallo-organic wear products is high. If a transition to damage-controlled wear occurs at low reactivity, oxides will predominate in the film. Extremely low reactivity for the severity level may lead to a further transition to surfacedamage wear with generation of metallic wear products. At high reactivity, the metallo-organic film becomes too thick, and the excess is removed by the sliding process at a rate dependent on severity. Although wear in the mild- and damage-controlled regimes is often termed “abrasive” or Copyright © 1997 CRC Press, LLC.

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“adhesive,” the bearing metal is a “corrosive” chemical reaction product in the boundary film. Note that reactivity A in Figure 4 gives smaller wear at intermediate severity than at higher or lower severities, while wear increases monotonically with severity over most of the reactivity range. Load-carrying capacity (scuff limit, etc.) transitions seem to depend on the thickness and physical properties of a reaction product film. These depend on the particular combination of reactants (i.e., bearing, lubricant and atmosphere materials) and the severity of the bearing system as a chemical reactor. The amount of time the film shears and reacts in the load-carrying zone reactor is included as an important part of severity. Consequently, the combination of severity variables that leads to the sudden increase in wear and friction (and often surface damage) depends on the pathway to transition. The effect of the pathway can be surprising. For example, Figure 5 compares the film thicknesses at run-in durations A through E in Figure 3. Note that, if film thickness alone determines the ability to operate at higher severity, this would explain the fairly widespread experience of being able to maximize load-carrying capacity of a bearing more rapidly by running in at an intermediate severity.

FIGURE 5 Operational severity effect on boundary film thickness. Table 3 summarizes most of what is known about the properties of boundary films. The films are largely amorphous and, to be effective, they must stay on the substrates and shear more easily. The physical state shown is that apparent for the outermost portion of the film.

BOUNDARY LUBRICANTS A wide variety of base oils is used because of their various combinations of properties.8 Hydrocarbon base oils, led by processed natural “mineral oils,” predominate because most of their properties are good to excellent and their cost is comparatively low. Ester and polyglycol base oil use is rapidly expanding because of some superior properties for certain applications.910 Boundary lubricated wear and friction characteristics of base oils vary with chemical type and with viscosity.11 Friction and wear properties and severity for transitions of base oils can be modified by adding small amounts of chemicals (at concentrations of a few parts per million to several percent). The molecules used for these purposes all contain surface-attracting and base-oil solubilizing portions Copyright © 1997 CRC Press, LLC.

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as illustrated in Figure 6 for hydrocarbon base oils. Boundary lubrication additives are classed as friction modifier (lubricity or oiliness), antiwear, or extreme pressure (EP). Friction modifier lubricity additives, such as oleic acid, lower the friction coefficient at low sliding velocity. This dampens friction-induced noise and can eliminate the stick-slip that occurs at very low sliding velocity when static friction exceeds sliding friction. Friction modifier additives also tend to increase the severity for the transition from mild to damage-controlled wear, but to decrease the severity for inducing surface-damage wear. Friction modifier molecules generally have long (at least 8 to 10 backbone atoms) flexible chains in the oil-solubilizing portion. Antiwear and EP additives usually reduce wear in the mild and damage-control wear regimes and increase the severity for the mild to damage-control wear transition. Antiwear additives may lower the load for the transition to surface-damage wear, but EP additives always increase the severity for this transition. Both antiwear and EP additives usually contain one or more of the elements sulfur, phosphorous, or chlorine as illustrated by the zinc diisopropyl dithiophosphate molecule in Figure 6. Both antiwear and EP additives are often used in a lubricant, and sometimes both functions are provided by a single molecule, such as a zinc dithiophosphate. Base oils, boundary lubrication additives, additives used for other purposes, and water and oxygen interact with both positive and negative effects on wear protection and transition severities.11

BOUNDARY LUBRICATION DETERMINATION Wear, friction, and failure analysis or prediction generally require an asperity interaction parameter calculation to determine the regime of lubrication. It also requires estimation of the likely boundary lubrication (sub-)regime, if in the boundary lubrication or mixed lubrication regime. INTERACTION PARAMETER CALCULATION Interaction parameter (IP) is the ratio of composite surface roughness to minimum hydrodynamic film thickness (i.e., (film thickness parameter)-1 = A-1). Its calculation requires determination of the composite roughness of the opposing bearing surfaces and calculation of the minimum hydrodynamic film thickness expected for smooth surfaces. Composite surface roughness is the Copyright © 1997 CRC Press, LLC.

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FIGURE 6 Examples of friction modifier and antiwear/EP additives for oil-based lubricants.11 root-mean square (RMS) surface roughness in which Rql and Rq2 are the RMS roughnesses of surfaces 1 and 2, respetively. Minimum fluid film thickness, hm, for hydrodynamic and elastohydrodynamic smooth surfaces can be calculated using equations and/or charts12,13 summarized in Chapter 58. The equations are quite accurate, provided that accurate lubricant viscosity properties are used and IP is less than about 1/2 ( Λ > ∼2 ). Accurate viscosity properties require accurate lubricant temperature, since viscosity is a strong function of temperature, and the viscosity-pressure coefficient for elastohydrodynamic film thickness calculation varies importantly with viscosity at atmospheric pressure (and consequently with temperature). The viscosity-pressure coefficient for elastohydrodynamic applications may be determined from equations and a chart given by Fein in Chapter 59, based upon Roelands.14 Note that the viscosity-pressure coefficient for elastohydrodynamic calculations is not the same coefficient appropriate for calculating viscosity at some specific pressure. Hence, care must be exercised before using coefficients reported in the literature. Approximate IP and A limits for the regimes of lubrication are shown in Table 4. A full thermal analysis of the bearing system or difficult measurement is usually required to accurately determine the appropriate bearing and/or lubricant temperature for bearing film thickness calculation. Winer and Cheng give methods of estimating the appropriate bearing temperature.13 Estimation of the boundary lubrication (sub)-regime may be accomplished by calculating the wear coefficient, observing the morphology of the wear surfaces, and observing the nature of the wear products. The following equations facilitate calculation of k for various contact conditions.15 For the ith surface in continuous contact

For the surface which periodically passes through the load-carrying zone with sliding parallel to the direction of motion (sweep direction) Copyright © 1997 CRC Press, LLC.

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If sliding and motion directions are different, the sliding velocity becomes the vector velocity difference and W/F becomes the product of the mean pressure and length of the load-carrying zone in the direction of surface motion.

Estimation of the amount of wear, Vi or di, that has occurred in a bearing is often challenging. Table 5 gives minimum detectable wear estimates for a number of common measurement methods. Measurement of a wear scar width or diameter is useful only when the same spot on one surface always carries the load, and the relative radius of curvature at that loadcarrying zone is sufficiently small (usually < ~100 mm). The most sensitive wear depth estimates can often be made by profilometry or, less precisely, by visual inspection of percent removal of surface finish marks (e.g., tool marks). Wear depth measurements or estimates can be integrated over the wearing area to give a wear volume.

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The following equations can be used to calculate wear volume from wear scar width for “point” and “line” contacts.15 Peterson15 also tabulates k vs. scar width for some standard bench wear tests.

BOUNDARY LUBRICATION ASSESSMENT FOR DESIGN AND PROBLEM ANALYSIS The likelihood of problems either with new pioneering designs or designs for extending the operating severity of existing systems can be estimated by means of five operation factors shown Copyright © 1997 CRC Press, LLC.

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in Table 6. They should be used to supplement existing bearing design and analysis techniques and experience with similar bearing systems. The factors also help solve problems of existing systems and development of simulation tests.

Each operating factor consists of the ratio of a severity measure for bearing system operation to a limitation for that measure. Magnitude and direction of the departure of the factor from unity indicates the degree to which operating severity is below or above the limitation imposed by the system configuration and materials or by the needs for system functioning. The five factors in Table 6 are described below. All factors may not be relevant to a given bearing system. 1. Surface interaction factor, as previously discussed, is the ratio of the minimum calculated hydrodynamic film thickness to the limitation imposed by the composite surface roughness. It indicates whether the system is operating in the full or partial boundary lubrication (i.e., mixed lubrication) regimes. If the latter, it estimates the departure from full bulk-lubricant fluid film surface separation. 2. Load factor estimates the limitation on maximum load imposed by the system configuration and materials. Commonly, it is the ratio of the mean pressure on the bearing to the pressure that begins to cause plastic deformation beneath the softer bearing material surface (often about 1/3 the indentation hardness at the bearing temperature). On occasion, the onset of asperity plasticity may be important. As a measure of relative asperity strength, the “plasticity index” may be used as an asperity load factor.16 Plasticity index

where s is the composite roughness, r is

the asperity tip radius, H is the indtation hardness of the softer material, and E’ is the effective elastic modulus for the two bearing surfaces. The hardness and effective modulus are at the bearing surface temperature. 3. The thermal factor estimates the bearing system approach to a thermal limitation. a. The measure may be a maximum surface temperature in a concentrated conjunction. The absolute temperature should be used in the factor and its limiting value. The limitation typically would be the scuff temperature limit for a gear or the bearing material softening temperature, or transformation temperature for a plain or thrust bearing.

b. PV, fPV, or PVT may also be used as measures of the thermal stress on a bearing. In these measures, P is the mean pressure on the load supporting zone, V is the sliding velocity, f is the friction coefficient, and T is the length of the load-carrying zone in the direction of sliding. 4. The wear factor is the radio of the dimensionless wear coefficient or of dimensional wear measures (e.g., volume of wear per unit of sliding distance and load). The limiting value

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is the maximum wear coefficient calculated to allow the bearing to function adequately for its intended life. 5. The friction factor is the ratio of the friction coefficients when too-high friction represents the limiting condition or the reciprocals of the friction coefficients when too-low friction is the limiting condition.

Table 7 lists the estimated ease of meeting the boundary lubrication limitations for each of the five operation factors. Table 8 gives some thermal factor limits compiled from literature assessments.17’18 Last, Table 9 shows typical wear coefficient ranges and their ease of achievement for ferrous metals. Limited experience indicates that the same ranges and degree of achievement difficulty apply to other bearing metals, but the effective lubricant additive packages frequently are different from those for ferrous metals.

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REFERENCES 1. Fein, R. S., A perspective on boundary lubrication, Ind. Eng. Chem. Fundam., 25(4), 518–524, 1986. 2. Fein, R. S., Boundary lubrication, Lubr. Eng., 47(12), 1005–1008, 1991. 3. ASME, Wear Control Handbook, Peterson, M. B. and Winer, W. O., Eds., American Society of Mechanical Engineers, New York, 1980. 4. Bowden, F. P. and Tabor, D., The Friction and Lubrication of Solids, Part I, Clarendon Press, Oxford, 1954. 5. Bowden, F. P. and Tabor, D., The Friction and Lubrication of Solids, Part II, Clarendon Press, Oxford, 1964. 6. Rabinowicz, E., Friction and Wear of Materials, Wiley, New York, 1965. 7. Rowe, C. N., Lubricated Wear, in CRC Handbook of Lubrication: Theory and Design, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1983. 8. Klamann, D., Lubricants and Related Products, Verlag Chemie, Weinheim, Germany, 1984. 9. Fein, R. S., Liquid lubricants, in ASM Handbook: Vol. 18, Friction, Lubrication, and Wear Technology, Blau, P., Ed., ASM International, Metals Park, OH, 1992, 81–88. 10. CRC Handbook of Lubrication: Monitoring, Materials, Synthetic Lubricants and Applications, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994. 11. Fein, R. S., Boundary lubrication, in CRC Handbook of L ubrication: Theory and Design, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1983, 49–68. 12. Hamrock, B. J., Fundamentals of Fluid Film Lubrication, McGraw-Hill, New York, 1993. 13. Winer, W. O. and Cheng, H. S., Film thickness, contact stress and surface temperature, in Wear Control Handbook, Peterson, M. B. and Winer, W. O., Eds., American Society of Mechanical Engineers, New York, 1980. 14. Roelands, C. J. A., Correlational Aspects of the Viscosity — Temperature-Pressure Relationship of Lubricating Oils, Doctors thesis, Technical Highschool of Delft, Druk. V.R.B., Groningen, Netherlands, 1966. 15. Peterson, M. B., Design considerations for effective wear control, in Wear Control Handbook, Peterson, M. B. and Winer, W. O., Eds., American Society of Mechanical Engineers, New York, 1980, 413–473. 16. Greenwood, J. A. and Williamson, J. B. P., Contact of nominally flat surfaces, Proc. R. Soc., Ser. A , 295, 300, 1966. 17. Drago, R. J., Comparative load capacity evaluation of CBN-finished gears, Gear Technology, 7(3), 8–16, 48, 1990. 18. Errichello, R., Friction, lubrication and wear of gears, in ASM Handbook: Vol. 18, Friction, Lubrication and Wear Technology, Blau, P., Ed., ASM International, Metals Park, OH, 1992, 535–545.

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Wear Problems — 46 Lubricated Symptoms and Prevention Douglas Godfrey

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BIBLIOGRAPHY Component failures, maintenance and repair, in Tribology Handbook, Neal, M. J., Ed., SAE, R137, 1995. ASM Metals Handbook, Vol. 10, Failure of sliding bearings, ASM, Metals Park, OH, 1975, 397–415. ASM Handbook, Vol. 18, Friction, lubrication and wear, Blau, P., Ed., ASM International, Metals Park, OH, 1992. Godfrey, D., Common wear problems related to lubricants and hydraulic fluids, Lubr. Eng., 43(2), 111–114, 1987. Jahanmir, S., On the wear mechanisms and the wear equations, Fundamentals of Tribology, Suh, N. P. and Saka, N., Eds., MIT Press, Cambridge, MA, 1980, 455–67. Kaufman, H. N., Sliding bearing damage, CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 477–494. Source Book on Wear Control Technology, Rigney, D. and Glaeser, W. A., Eds., ASM, Metals Park, OH, 1978. Tallian, T. E., Failure analysis for Hertz contact machine elements, ASME Press, New York, 1992. Wear Control Handbook, Peterson, M. B. and Winer, W., Eds., ASME, 1358 pages, 1980. (See pages 1143–1201 for Glossary of Tribological Definitions by the Organization For Economic Cooperation and Development). Bulletins of engine, industrial equipment, components, and lubricant manufacturers.

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VI Material Properties

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PROPERTIES OF 47 TYPICAL SLIDING CONTACT MATERIALS A. W. Ruff

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ELEMENT BEARING 48 ROLLING MATERIALS Charles A. Moyer Rolling element bearing materials started with modification of tool steel in Europe about 1900, leading to AISI 52100 and derivatives that have since been used for all types of commercial and automotive rolling bearings. In the U.S. AISI 52100 was the primary choice for industrial ball bearings, spherical bearings, and other types depending on the bearing manufacturer. Carburizing grade steels were also developed early for tapered roller bearings and other heavily loaded bearing types that benefited from the tougher core materials and case compressive residual stresses that developed during the carburizing process.1 Both the popular through-hardened steels (i.e., AISI 52100) and carburized steels (i.e., AISI 8620) contain about 1.5 to 6% alloy content, 0.6 to 1.1% C (approximate carbon content range), and the rest iron. They have been fairly cost effective, with sufficient loading and fatigue strength to perform well. As the inclusion content size and number have been reduced in recent times, the fatigue strength has increased dramatically for these steels. As rolling element bearings were used beyond automotive and industrial applications, other materials were developed. Examples are steels with higher alloy content (10 to 15% or more) for high temperature and better corrosion resistance (e.g., 440C) or ceramics (e.g., silicon nitride or alumina) for higher strength, higher stress limits, and reduced centrifugal force effects. Tables 1 and 22 provide compositions of standard low-alloy content-bearing steels and Table 32 core properties of some of these steels. Tables 4 and 52 list other materials that have been seriously considered for use in rolling element bearings. Some of the special materials have been tested only in laboratory rigs. Hoo3 provides a good summary of several bearing material test rigs with results on various materials. Silicon nitride was the first ceramic that has been developed to the point of special commercial use. Costs of bearings from these special materials can be 5 to 20 times the cost of standard industrial grade bearings because of higher alloy content or special processing. Because of their exceptional capabilities, however, these materials will permit rolling element bearings to be used in extremely hostile environments and much wider operating temperatures or adverse lubrication and stress conditions. Zaretsky4 has provided summaries of typical physical and thermal properties of some special materials as given in Table 6. For standard bearing steels, Rockwell C hardness is usually in the range of 58 to 63 Rc. An approximate temperature limit is 150°C (300°F) for standard steels based on the usual temper levels. An elastic modulus of 207 GPa (30 million psi) and Poisson’s ratio of 0.3 can be used in usual calculations for contact stresses and film thickness equations. The Weibull slope has often been considered a material “modulus” and in a variety of fatigue tests it has had values ranging from just less than 1.0 to over 3.0. If specific values are needed, it is recommended that they be obtained directly from the bearing manufacturer. The same is true for the other material properties; reliable values and their range of applicability can be obtained from the specific bearing steel maker. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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Figure I5 gives the composite fracture toughness for various carburizing and homogeneous high-carbon steels. Very often fracture toughness is a critical property that should be considered in bearing material selection. Table 7 is a summary of bearing steels ordered by operating temperature limits (compiled by E. E. Pfaffenberger for STLE Rolling Bearing Fundamentals and Damage Analyses courses). Information on rolling bearing cage or separator materials is in Table 8,6 covering a range of lubrication and temperature conditions. For in-depth reviews of bearing materials, their processing, and their failures, see Anderson et al.,7 the chapter on “Materials and Processing” and Moyer and Zaretsky.8

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FIGURE 1 Composite fracture toughness of carburizing and homogeneous high-carbon steels in slow bending. Case depth is 0.76 to 0.89 mm (0.030 to 0.035 in.) to 0.50% C level. Shaded areas indicate range of K values for cracks originating in core. Cross-hatched areas indicate range of K values for cracks originating in case. Charpy-sized specimens were carburized, hardened, tempered, and precracked to several depths in case and core regions before testing. As cracks progress inward, the fracture resistance of carburized composites improves significantly. (From Jatczak, C. R, Met. Prog., April 1978. With permission.) Copyright © 1997 CRC Press, LLC.

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REFERENCES 1. Rowland, E. S., Resistance of materials to rolling loads, in Handbook of Mechanical Wear, Frettage, Pitting, Cavitation and Corrosion, Lipson, C. and Colwell, L. V., Eds., University of Michigan Press, Ann Arbor, MI, 1961, 108–130. 2. Burrier, H. I., Bearing steels, in Metals Handbook, Vol. 1, 10th ed., Properties and Selection, ASM International, Metals Park, OH, 380–388, 1990. 3. Hoo, J. J., Ed., Rolling Contact Fatigue Testing of Bearing Steels, ASTM STP-771, American Society for Testing and Materials, Philadelphia, 1982. 4. Zaretsky, E. V., Ceramic bearings for use in gas turbine engines, in J. Eng. Gas Turbines and Power, 111(1), 146–157, 1989. 5. Jatczak, C. F., Specialty carburizing steels for high temperature service, Met. Prog., April 1978. 6. Waterman, N. A. and Ashby, M. F., Eds., CRC-Elsevier Materials Selector, Vol. 1, 364, 1991. 7. Anderson, W. J., Bamberger, E. N., Poole, W. E., Thom, R. L., and Zaretsky, E. V, Materials and processing, STLE Life Factors for Rolling Bearings, Zaretsky, E. V., Ed., STLE Pub. SP-34, Society of Tribologists and Lubrication Engineers, Park Ridge, IL, 1992, 71–128. 8. Moyer, C. A. and Zaretsky, E. V, Failure modes related to bearing life, STLE Life Factors for Rolling Bearings, STLE Pub. SP-34, Zaretsky, E. V, Ed., Society of Tribiologists and Lubrication Engineers, Park Ridge, IL, 1992, 47–69.

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49 Oil Film Bearing Materials George R. Kingsbury CONTENTS Introduction.......................................................................................................................................503 Distinctive Property Requirements................................................................................................503 Characteristics of Metallic Bearing Material Systems..................................................................503 Individual Bearing Alloys.................................................................................................................504 Other Metallic and Nonmetallic Bearing Materials......................................................................504 Defining Terms..................................................................................................................................505 For Further Information..................................................................................................................505

INTRODUCTION Commercially practical oil film bearing materials are predominantly but not exclusively metallic, and in large part are comprised of two or more separate bonded layers. The information in this section has therefore been organized to reflect the performance of these materials in the context of both single-layer and multilayer systems, as well as to provide relevant data on compositions and properties of the materials that form the separate layers. A. Table 1: Distinctive property requirements B. Tables 2 to 4: Characteristics of one-, two-, and three-layer metallic systems C. Tables 5 to 13: Composition, form, functions, and properties of individual bearing alloys D. Tables 14 to 15: Characteristics and properties of other metallic and nonmetallic bearing materials

DISTINCTIVE PROPERTY REQUIREMENTS Of the six bearing material attributes listed in Table 1, only hardness and compressive strength are measurable directly by conventional laboratory test methods. Special chemical and highspeed dynamic mechanical test methods and equipment have been developed to evaluate the other attributes and their interactions.

CHARACTERISTICS OF METALLIC BEARING MATERIAL SYSTEMS In general, fatigue resistance and load capacity in Tables 2 to 4 vary directly with hardness and compressive strength; whereas conformability and embeddability vary inversely. Compatibility ratings refer specifically to adhesive wear processes with steel counterfaces. These ratings have no necessary relationship to material hardness or strength. Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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INDIVIDUAL BEARING ALLOYS Strength retention at elevated temperatures in Tables 5 to 13 is especially important in internal combustion engine and gas turbine bearing applications. Materials selection for these kinds of service must take this dependency into account. The relatively rapid weakening of zinc, lead, and tin-base alloys with increasing temperature, illustrated in Figure 1, restricts their use in such applications.

FIGURE 1 Strength retention at elevated temperatures for selected bearing alloys, (a) Copperbase alloys, (b) Aluminum-base alloys, (c) Zinc-base alloys, (d) Lead-base alloys and tin-base alloys. (From Kingsbury, G.R., ASM Handbook, Vol. 18, ASM International, leveland, OH, 1992, 745. With Permission.)

OTHER METALLIC AND NONMETALLIC BEARING MATERIALS The carbon and polymeric materials in Tables 14 and 15 are most frequently used for “dry rubbing” bearings that can operate without oil lubrication. In general, however, they can operate under more extreme conditions of loading and speed if oil is present. They are included here on that basis, with PV ratings that correspond to oil-lubricated operation. Copyright © 1997 CRC Press, LLC.

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DEFINING TERMS Compatibility — The antiwelding and antiscoring characteristics of a bearing material when operated with a given mating material. Some potential for scoring and seizure exists under all boundary and thin-film lubrication conditions. Conformability — The ability of a bearing material to yield plastically and to compensate for small misalignments, variations in the shape of the journal, or of the bearing-housing bore. Embeddability — The ability of a bearing material to embed dirt or other foreign particles and thus prevent them from scoring and wearing journal and bearing surfaces. Fatigue strength — The ability of a bearing material to function under cyclic loading conditions without developing cracks or surface pits. Hardness/compressive strength — The ability of a bearing material to resist deformation under high unit compressive loads. Hardness is conventionally measured by indentation hardness testing. Compressive strength may be determined directly by compression testing of material specimens or by inference from indentation hardness tests. Corrosion resistance — The ability of a bearing material to withstand chemical attack by uninhibited or contaminated lubricating oils.

FOR FURTHER INFORMATION A good treatment of the underlying principles of oil film bearing design and operation is presented in “Friction and Wear of Sliding Bearings” by Ron Pike and J.M. Conway-Jones, in ASM Handbook, vol. 18, pp. 515–521 (1992). “Friction and Wear of Sliding Bearing Materials” by this author, pp. 741–757 of the same volume, gives a fairly comprehensive summary of current sliding bearing materials technology. Good introductory treatments of bearing materials technology are presented in “Sliding Bearing Materials” by A.O.DeHart, in CRC Handbook of Lubrication, vol. 2, pp. 463–476 (E.R. Booser, Ed. —1984); and in “Bearing Materials” by E.R. Booser, in Kirk-Othmer Encyclopedia of Chemical Technology, 4th ed. (1992). Bearings —A Tribology Handbook (M.J. Neale, Ed.), published in 1993 by The Society of Automotive Engineers (SAE), presents practical guidance on the selection and design of both rolling contact and sliding bearings, including both oil-lubricated and dry rubbing operation.

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Properties of Gear 50 Mechanical Materials Lewis Rosado Material selection is an important process when designing or choosing a gear for any given application. The designer or user must consider not only factors such as material availability, cost, load-carrying capacity, manufacturing requirements, gear size, and weight, but also operating parameters such as temperature, load, speed, type of lubrication, and expected reliability. Some of the most critical material properties which influence gear performance include tooth surface hardness; core fracture toughness and bending strength; and fatigue, impact, and wear resistance. Corrosion resistance is also important, specifically when the service requirements include a corrosive working medium or a highly oxidizing environment. While gear materials range through the nonferrous metals, sintered powder metals, and several plastics, ferrous metals continue to be the most widely used because of their high strength, low cost, and heat treatment response. Ferrous materials used for gearing include cast irons, plain carbon steels, and alloy steels. For moderate load applications, cast iron provides excellent damping properties under dynamic conditions, good sliding and wear characteristics, and minimum machining costs. The plain carbon steels are relatively easy to machine, inexpensive, can be hardened, and are used in power gearing with moderate load reatings despite their poor corrosion resistance. The low, medium, and high alloy steels offer the widest range of mechanical properties, provide the highest strength and durability among the gear materials, are the most versatile, and can be processed to withstand severe power requirements. Nonferrous gears are made from alloys of copper, aluminum, and zinc. However, the copper alloys, particularly the bronzes, account for most of the nonferrous gear materials, mainly because of their “wear resistance” characteristics for withstanding a high sliding velocity with a steel worm gear. For example, most worm gearsets use a carburized and ground steel worm and a hobbed bronze wheel because hardened, smooth steel sliding on ductile, lubricated bronze performs well. In contrast, a stainless steel worm sliding on a stainless wheel is likely to scuff even at low loads. Plastic gear materials are mostly based on acetate and nylon resins; polyimide gears are also used for extreme temperature conditions or when self-lubricating gears are required. Characteristics which make plastic gears attractive for mild operating conditions at low load include low cost, low noise during operation, and inherent wear resistance. Plastic gears also require minimum or no lubrication. Limitations include low strength and temperature resistance, and poor dimensional stability and accuracy. Recent advances made in gear material technology have primarily been a result of the need for improved helicopter transmission systems and aircraft gas turbine engine mechanical components. Improvements is steel metallurgy and processing have enabled operation at higher temperatures, speeds, and loads, while providing substantial increases in reliability. Several materials which were originally developed as high-temperature rolling element bearing materials have been evaluated as potential gear materials (Townsend and Zaretsky, 1974; Townsend, Bamberger, and Zaretsky, 1976; Townsend, Parker, and Zaretsky, 1979; Townsend, 1985; Townsend and Bamberger, 1991). Some Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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of the most promising high hot-hardness materials include vacuum-induction melted, vacuumarc remelted M50 NiL, CBS 600, CBS 1000M, X-53, and Vasco X-2M. Other materials which have been recently developed as corrosion resistant bearing materials, namely, Pyrowear* 675 chrome tool steel (Pfaffenberger and Tarratini, 1993; Wert, 1994) and Cronidur** 30 (a nitrogen-alloyed martensitic steel) (Trojahn, 1992), may also be suitable as high temperature, corrosion-resistant gear materials. Tables 1 to 16 list the chemical composition and some commonly cited mechanical properties for a number of gear materials.

FOR FURTHER INFORMATION Gear standards, which also include gear material standards, have been developed and sponsored by the American Gear Manufacturers Association (AGMA). For further information on gear standards contact the AGMA at 1500 King St., Suite 201, Alexandria, VA 22314 or by phone at (703) 684-0211.

* Tradename of Carpenter Steel Co., Reading, Pennsylvania. ** Tradename of FAG Bearing Ltd., Stratford, Ont.

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Gearing by Townsend and Zaretsky (NASA RP-1152, 1985) provides additional information on the processing, heat treatment, and properties of gear materials as well as gear design theory and historical aspects of gearing. Status of Understanding for Gear Materials by Townsend (NASA CO-2300, p. 795, 1983) offers an excellent summary of the practical aspects of gear materials and also includes some material property data.

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REFERENCES Dudley, D.W., Practical Gear Design, 1st ed., McGraw-Hill, New York, 1954. Fopiano, P.J., Krzanowski, J.E., and Crawford, G.M., Direction of R&D and current status of understanding of advanced gear steels, AGARD Conf. Proc., No. 394, 1985, 9.1–9.15. Hamrock, B.J., Fundamentals of Fluid Film Lubrication, McGraw-Hill, New York, 1994. Handbook of Gears, D190, Stock Drive Products, Sterling Instruments, New Hyde Park, NY, 1992, p. T93. Jackson, E.G., Muench, C.F., Rowe, E.H., and Scott, E.H., Evaluation of Alloys for High Temperature Gear Applications, WADC TR-58-546, 1958. Metals Handbook: Properties and Selection: Irons and Steels, Vol. 1, 9th ed., American Society for Metals, Metals Park, OH, 1978. Metals Handbook: Properties and Selection: Nonferrous Alloys and Pure Metals, Vol. 2, 9th ed., American Society for Metals, Metals Park, OH, 1979. Pfaffenberger, E.E. and Tarrantini, P., High Temperature Corrosion Resistant Bearing Steel Development, AIAA 93-2000, 1993. Shigley, J.E., Mechanical Engineering Design, 2nd ed., McGraw-Hill, New York, 1972. Townsend, D.P., Dudley’s Gear Handbook, 3rd ed., McGraw-Hill, New York, 1980. Townsend, D.P., Surface Fatigue Life and Failure Characteristics of EX-53, CBS 1000M, and AISI 9310 Gear Materials, NASA TP-2513, 1985. Townsend, D.P. and Bamberger, E.N., Surface fatigue life of M50NiL and AISI 9310 gears and rolling-contact bars, J. Propul., 4(7), 642–649, 1991. Townsend, D.P., Bamberger, E.N., and Zaretsky, E.V., A life study of Ausforged, standard forged, and standard machined AISI M50 spur gears, Trans. ASME, 20(75), 418–425, 1976. Townsend, D.P., Parker, R.J., and Zaretsky, E.V., Evaluation of CBS 600 Carburized Steel as a Gear Material, NASA TP-1390, 1979. Townsend, D.P. and Zaretsky, E.V., A life study of AISI M50 and super nitralloy spur gears with and without tip relief, J. Lubr. Tech., 38(73), 583–589, 1974. Trojahn, W., High Nitrogen Martensitic Steels — A New Family of Martensitic Corrosion Resistant Steels for Improved Aerospace Bearing Performance, ASME Pub., 92-GT-338, ASME, New York, 1992.

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51 Wear-Resistant Hard Materials William A. Glaeser CONTENTS Introduction....................................................................................................................................540 Hard Elements................................................................................................................................541 Hard Facing Materials....................................................................................................................541 Intermetallics..................................................................................................................................542 Cemented Carbides........................................................................................................................544 Glasses.............................................................................................................................................544 Diamond and Diamond Coatings................................................................................................545 White Cast Iron..............................................................................................................................545 References.......................................................................................................................................546

INTRODUCTION One guiding principle used in the selection of wear-resistant materials is that high hardness means good wear resistance. There are some exceptions to this general rule: hard, brittle materials will tend to spall and accelerate wear because of low fracture toughness. Wear-resistant hard materials, generally of greater hardness than heat-treated steel and often of low fracture toughness, are divided into the following distinct classes in this section: hard facing (Stellites); hard elements (tungsten, chromium, iridum); intermetallics (titanium nitride); cemented carbides (tungsten carbide); glass; diamond; and white cast iron. Aside from hard chromium plate, this section will not cover coatings such as plasma spray, PVD and CVD deposits, ion plating, and ion implantation. Where possible, abrasion, erosion, and wear-resistance data will be provided for the materials listed. The data will be from standard ASTM tests such as the rubber wheel abrasion test, block-on-ring, pin-on-disk, and air jet erosion tests. In addition to wear properties, mechanical, physical, and thermal properties will be provided. When dry sliding friction creates sufficient surface heating, especially at localized spots, a rapid rise and fall of temperature will produce spalling and an acceleration of wear. Thermoelastic theory suggests a thermal gradient in a material with low thermal conductivity, and high elastic modulus makes it more prone to thermal shock and fracture. For the condition of sudden rapid rise in local surface temperature, the thermal shock resistance of a material can be estimated from the following:1

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Where: TSR σf k α E

= thermal shock resistance = tensile fracture stress, MN2 = thermal conductivity, Wm-1degm-1 = thermal expansion coeff., 10m-6degm-1 = Young’s modulus, MN/m2

The lower the TSR number, the higher the sensitivity to thermal shock fracture as, for instance, with Al2O3 with TSR = 1.4 as compared with much lower sensitivity in tool steel with TSR = 57.

HARD ELEMENTS Several high-melting point metallic elements are listed in Table 1 which have inherent high hardness. It should be cautioned that values found in the tables in this chapter can vary by as much as 15%, depending upon the source of the data. Tungsten and osmium achieve their maximum hardness by heavy working. Chromium becomes extremely hard and wear resistant when used as an electroplated coating. Chromium plate in the industrial hard plate condition is in its hardest condition.

Iridum, osmium, and rhodium are part of the platinum group and are considered precious metals. They have been used as pen nibs and electrical contacts. Chromium is used as an abrasion and corrosion resistant coating for shafts and seal rotors. Tungsten has been used as pins in dot matrix printers.

HARD FACING MATERIALS Cobalt, nickel, or iron-base alloys with high hardness and excellent abrasion resistance are used as weldments applied over steel surfaces to protect them from abrasive wear. They are used in the mining industry on drag line buckets, repair of teeth on power shovel buckets, in ore crushing machinery, pump impellers, etc. Because the alloys can be applied by welding techniques, they are convenient for field repairs. The hard facing materials, such as those in Table 2, have excellent corrosion resistance and can be used effectively at high temperatures. They have been used in bulk form as components in 600(F water nuclear reactors, sea water face seals, engine valve faces, and seals in the petrochemcial industry. Literature on wear and mechanical properties of hard facing materials has diminished in volume considerably in the last decade. Probably the best reference is the 1974 proceedings of the symposium on Meterials for the Mining Industry.4

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Since the cobalt-base satellites are used in weld coatings applied by oxyacetylene and gas tungsten are methods and as castings, their properties can vary depending on the method of fabrication. The satellites are large chrome carbides in cobalt matrix. The Triballoys are high-temperature alloys, either cobalt base or nickel base with molybdenum, chromium, and silicon, and low carbon content. These alloys develop their hardness by an intermetallic Laves phase rather than massive carbides. Therefore, they are more resistant to impact conditions that can fracture carbides. The abrasion resistance of Triballoys is equivalent to satellites. The composition of these hard facing alloys are found in Table 3. Table 4 shows their abrasion resistances, as determined by the ASTM low-stress rubber wheel abrasion test.8

INTERMETALLICS When some metals are alloyed, they solidify with intermetallic phases. These phases are chemical reaction products. Chromium carbides in alloyed white cast iron are an example. Carbides are also intermetallics, and they have been used in the cemented form with cobalt or nickel as binder. Usually, these intermetallics are very hard and abrasion resistant. Producing intermetallics as singlephase materials has been of growing importance. Cubic boron nitride is a hard intermetallic (7300–10,000 HV)9 that is close to diamond in hardness and is used as a cutting tool material. The triballoy alloys used as hard facing alloys were Copyright © 1997 CRC Press, LLC.

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developed in an attempt to achieve tough, high-temperature wear-resistant alloys without massive carbides in the structure. Many other intermetallics are very hard, but are not practical because of extreme brittleness and the difficulty of fabricating them into components. In recent years, however, intermetallic materials have begun to be used in industrial electronic devices. One such material is nickel aluminide, Ni3Al, which has the unusual property of increasing in yield strength with temperature as shown in Figure 1. When modified by the addition of boron to increase its toughness, this alloy has shown dry wear characteristics similar to dual-phase steel. Lubrication improves the wear resistance considerably. Although this alloy is not exceptional in its wear behavior, it represents a whole class of hard materials that may develop for practical use. Other intermetallics with future potential include Ti3Al, TiN, and AIN.

FIGURE 1 Yield strength of Ni3Al as a function of temperature. The alloys Hastelloy-X and 316 stainless have been included for comparison.10 (From Blau, P. and Devore, C. E., Tribol. Int., 23(4), 226–234, 1990. With permission.)

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CEMENTED CARBIDES Carbides of the transition metals shown in Table 5 are very hard and abrasion resistant. They are found in steels as hardening agents. Tungsten carbide and titanium carbide grains are used in cemented form in a binder or matrix of cobalt or nickel. These materials are used in tool bits and rock drills. Cemented tungsten carbide is produced by liquid-phase sintering of cobalt and tungsten carbide particles. The carbides provide the abrasion-resistant constituent and the cobalt provides the fracture toughness needed for tool applications. The wear properties of some carbides are summarized in Table 6.

GLASSES Glasses such as shown in Table 7 are solid metal oxides without long-range crystal habit. As they are cooled from the liquid state, their viscosity increases and they harden without any sudden change in density. They posses a glass transition temperature, however, and their thermal expansion coefficient changes at that temperature. The expansion coefficient is larger in the melt condition.

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Glasses are transparent and very brittle. While unsuitable for Hertzian contact or bending, glasses have been used for thread guides in yarn spinning machines. Glass fibers are used as reinforcement for plastics and composites in which the glass fiber reinforcement reduces wear rate of the composite at the expense of an abrasive effect which can wear tooling. Pyroceram is made by precipitating small amounts of metal crystals in glass. Heat treating then causes glass crystals to nucleate and grow around the metal particles. Pyroceram is translucent and, with high thermal shock resistance, has been used in bearing and seal applications.

DIAMOND AND DIAMOND COATINGS Diamond, the hardest material and very abrasion resistant, has been used as a griding compound, in polycrystalline composites as tool materials, and in wear-resistant coatings. Although its hardness and high thermal conductivity make the diamond ideal for wear-resistant applications, its surface converts to graphite and its wear rate increases if frictional heating becomes excessive. In addition, when used to cut ferrous materials, it tends to combine at the surface and form iron carbides. Cubic boron nitride, almost as hard, is used instead of diamond under these circumstances.

In recent years, diamond and diamond-like coating processes have been developed to the point where they are becoming practical in providing a surface which has the hardness of diamond without the expense of bulk diamond. There are still problems in getting diamond coatings to adhere to engineering alloys, however.

WHITE CAST IRON Hard, abrasion-resistant cast irons constitute high-chromium iron alloys. The range of compositions varies between 6 and 35 wt % chromium and 2 to 6% carbon. Compositions of

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some commercial white cast irons are shown in Table 9; Table 10 gives some of the mechanical properties. The abrasion-resistant irons are heat treated to achieve their high hardness. High chromium irons are used in the mining industry for metal-to-ore and –earth contact conditions, such as bucket teeth for power shovels. These alloys are used as hard facing, applied by welding. Since white irons are brittle, care must be taken in selecting alloys of suitable fracture toughness.

REFERENCES 1. Waytt, O. and Dew-Hughes, D., Metals, Ceramics and Polymers, Cambridge University Press, London, 1974, 229. 2. Smithells, C., Metals Reference Book, Butterworths, London, 1955. 3. Metals Handbook, 1984 ed., American Society for Metals, Metals Park, OH, 20–21. 4. Barr, R. Q., in Symp. On Materials for the Mining Industry, Vail CO, Climax Molybdenum Company, July 30 and 31, 1974. 5. Alloy Digest, Engineering Publications, Orange, NJ, [monthly]. 6. Stellite cobalt base rods, electrodes, and wires, Cabot Corp., Kokomo, IN, 1982. 7. Triballoy Wear Resistant Intermetallic Meterials, Cabot, Corp., Kokomo, IN, 1980. 8. ASTM Standard G 65–81, Standard Practice for Conducting Dry Sand/Rubber Wheel Abrasion Tests, ASTM, Philadelphia, 1981. 9. Westbrook, J. H. and Fischer, R. C., Intermetallic Compounds, John Wiley & sons, New York, 1995. 10. Balu, P. and Devore, C. E., Tribol. Int., 23(4), 226–234, 1990. 11. Wyatt, O. and Dew-Hughes, D., Metals, Ceramics and Polymers, Cambridge University Press, London, 1974, 258–268. 12. Hochman, R. F., Surface modification, Advanced Materials and Processes, American Society for Metals, Metals Park, OH, Jan. 1, 1995, 29–30. 13. Crook, P. and Farmer, H. N. Friction and wear of hardfacing alloys, ASM Handbook, Vol. 2, American Society for Metals, Metals Park, OH, 1993, 75–76. 14. Biesburg, D.E. and Borik, F., in Optimizing abrasion resistance and toughness in steels and irons for the mining industry, Symp. Proc. Materials for the Mining Industry, Val, CO, Climax Molybdenum Co., July 30 and 31, 1974, 26–28.

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Friction, Wear, and PV Limits of 52 Polymers and Their Composites Thierry A. Blanchet CONTENTS Introduction....................................................................................................................................547 Tribological Properties of Polymers............................................................................................548 Acknowledgment...........................................................................................................................550 References.......................................................................................................................................560

INTRODUCTION Table 1 provides representative values describing the tribological behavior of polymers and composites in dry, unlubricated sliding contact with finished metallic surfaces under ambient room temperature conditions. Data are provided for static and kinetic coefficient of friction, steady-state rate of wear (expressed as volume loss per distance slid, per normal load), and the limiting product of apparent contact pressure P and sliding velocity V under which the material may be safely employed. (When multiplied by the friction coefficient, the product PV dictates the rate of frictional heat dissipation per unit area, and therefore the contact temperatures at the sliding interface. Since for any given material critical temperatures exist, a limiting PV therefore exists). The effects of addition of a single hard particulate or solid lubricant filler to composites of these polymer matrices are also given. Loading levels of filler materials are either denoted by volume percent (v%) or weight percent (w%). Though more than one filler type is often employed, space does not permit the additional listing here of the numerous composites of matrix and two or more filler materials. The tabulated data originate from a broad variety of sources, including national laboratories, universities, and corporations, and were collected from publications in peer-reviewed professional journals, conference proceedings, and product bulletins. In most cases an effort was made to present single representative or average values for each composite system. When separate investigators reported very different behavior for very similar materials, both sets of data are presented. Such variations may result from slight differences in test materials, or from system differences involving counterface roughness and chemistry, load, spped, and contact kinematics, and atmosphere and temperature. Slight variations in these materials or sliding system attributes may lead to significant tribological differences, as illustrated in the following examples.

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TRIBOLOGICAL PROPERTIES OF POLYMERS SLIDING SPEED As polymers are viscoelastic materials and can be thought of as possessing characteristic stress relaxation times, sliding speed can be likened to an imposed strain rate. Therefore kinetic friction coefficient usually increases with increasing sliding speed.29 It may be for this reason that so many of the kinetic coefficients of friction reported from Reference 1, for example, exceed the static coefficients of friction, as kinetic friction in this case was measured at a speed of 0.25 m/s. (Kinetic friction at lower speeds is quite likely often less than the reported static friction.) At extremely high speeds, kinetic friction coefficient may eventually decrease with increasing speed from frictional heating and softening.29 As friction force often activates wear mechanisms, Blanchet and Kennedy25 have noted a transition in wear rate of PTFE from 10*l0-15 m3/Nm to 500*10-15 m3/Nm upon a small increase in sliding speed from 8 to 10 mm/s at room temperature. PV limit can also change significantly with sliding speed. For example, Theberge19 has shown that the PV limit of a polystyrene/silicone composite, 0.32*106 N/ms at 0.5 m/s, becomes as low as 0.035*106 N/ms as speed is increased to 5 m/s. PV limits within Table 1 were generally determined at a speed of 0.5 m/s. TEMPERATURE Relaxation time of polymers decreases with increasing temperature.52 As a result, decreases in ambient temperature often have effects equivalent to the increases in sliding speed.25,29,40 As PV limits are based upon the attainment of a critical temperature at the sliding interface, increases in ambient temperature will result in decreases in PV limit. Above the melt temperature, molecules have adequate thermal energy to move about freely as a viscous fluid. As temperature decreases below the melt temperature Tm, semicrystalline polymers will develop regions (crystallites) of three-dimensional order with a corresponding sharp decrease in free volume. The lack of crystallization in amorphous polymers having structure not conducive to ordering instead experience gradual reductions in free volume with decreasing thermal energy. Equally important property changes take place at the glass transition temperature. The disordered regions of semicrystalline polymers (as well as the entirety of amorphous polymers) lose sufficient thermal energy to become glassy below Tg, and the polymer therefore transforms from being rubber like and ductile to rigid and brittle. NORMAL LOAD Due to the lack of proportionality between real area of contact and normal load for polymers, the coefficient of friction in sliding contacts with smooth counterfaces tends to decrease with increasing load. For example, Janczak et al.41 have shown the coefficient of sliding friction of an EPDM composite rubber to decrease from µ = 1.1 to half that value as apparent contact pressure is increased from 0.1 to 0.6MPa. CONTACT KINEMATICS Entrapment of debris within a sliding contact can greatly affect its subsequent wear behavior. Likelihood of either entrapment or ejection of debris depends greatly on the type of sliding motion and on the degree of mutual overlap existing between the two contacting bodies. Debris entrapment becomes particularly likely in small amplitude oscillatory contacts. Abarou et al.,28 for example, have shown the wear rate of a PA 6/6 composite to increase 50-fold as mutual overlap coefficient MOC was increased from 0.33 to 0.8.

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COUNTERFACE ROUGHNESS Counterface surface roughness greatly affects the relative degree of adhesive and abrasive contributions to resultant friction and wear behavior of polymers. Generally, a minimum will exist. For example, at a surface roughness of Ra = 0.1 µm, the friction coefficient of HDPE will be less than µ= 0.2. As surface roughness is either decreased to Ra = 0.01 µm or increased to Ra = 1.0 µm, the coefficient of friction can exceed µ = 0.3. For LDPE, wear rate experiences a minimum of less than 10*10-15 m3/Nm at Ra = 0.025 µm. As roughness is decreased from this value to Ra = 0.01 µm, wear rate increases to 180*10-15 m3/Nm, while as roughness is increased to Ra = 1.3 µm, wear rate increases to 500*10-15 m3/Nm.30 Data tabulated here are from tests utilizing finished (polished or lapped) counterfaces, and data obtained under abrasive wear conditions utilizing abrasive papers, gauzes, or other rough counterface materials have been omitted. TRANSIENT RUN-IN The wear rates, as reported, are understood to be steady-state values. A very different rate of wear may exist at the initiation of sliding, gradually approaching a roughly constant value as steady-state conditions are attained after a few hundred meters of sliding. Generally, wear rate during this transient run-in period will be greater than that attained during steady-state. However, in instances involving counterfaces that are initially very smooth, wear rate may instead increase with time and level at a higher rate of steady-state wear as sliding roughens the counterface.30 Relative importance of transient and steady-state wear behavior will depend specifically upon the application and the anticipated sliding distance. COUNTERFACE SURFACE CHEMISTRY/ATMOSPHERE Friction and wear behavior can also be affected by interactions occurring between the polymer body and the counterface surface. Surface films which may exist and reform upon the counterface depend upon the counterface material as well as the atmosphere or environment in which sliding occurs. For similar surface roughnesses, the wear rate of a 10% graphite powder-filled PTFE increased over 30-fold if a steel counterface was replaced by aluminum.1 Data tabulated here are primarily from sliding contacts involving steel counterfaces under ambient air atmospheres. MOLECULAR WEIGHT/CRYSTALLINITY Hu and Eiss’s study of PTFE31 illustrates the potential effect of molecular weight and crystallinity on polymer wear. For a PTFE of lower crystallinity (0.35), particulates may no longer be well dispersed. The increasing occurrence of weak particle/particle interfaces may eventually result in increasing wear rate with increasing volume fraction,44 deviating from this inverse rule-of-mixtures behavior. For each data set in Table 1, a literature reference is provided. Since wide variation from the tabulated values exist, the reader is strongly advised to consult the corresponding reference and compare test conditions before further considering application of any material listed within this tabulation. Furthermore, it is assumed that the reader has available from other sources the compressive load limits and useful temperature ranges of the various polymer materials, as well as other nontribological characteristics which may affect their successful implementation in bearings applications.

ACKNOWLEDGMENT The author acknowledges support of the National Science Foundation Young Investigator Program under Grant No. CMS-9457596 during the preparation of this contribution. Any opinions, findings, and conclusions or recommendations expressed in this material are those of the author and do not necessarily reflect the views of the National Science Foundation.

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REFERENCES 1. Lubricomp Internally Lubricated Reinforced Thermoplastics and Fluoropolymer Composites, Bull. 254–688, LNP Engineering Plastics, Malvern, PA, 1988. 2. Lubricomp: A Guide to LNP’s Internally Lubricated Thermoplastics, Bull. 254–1094, LNP Engineering Plastics, Malvern, PA, 1994. 3. Thermocomp Carbon Fiber Reinforced Thermoplastics and Melt Processible Fluoropolymer Composites, Bull. 222–687, LNP Engineering Plastics, Malvern, PA, 1987. 4. Fluorocomp Composites, Bulletin 106–686, LNP Engineering Plastics, Malvern, PA, 1986. 5. Fluorocomp Filled PTFE Compounds, PD 109–880, ICI Fluoropolymers, Exton, PA, 1980. 6. Bronze-filled PTFE Fluorocomp Composites, Bull. 102–1288, ICI Fluoropolymers, Wilmington, DE, 1988. 7. Victrex PEEK — The High Temperature Engineering Thermoplastic, Literature Reference VK10/0690, ICI Advanced Materials, Exton, PA, 1990. 8. Vespel, Tech. Bull., Du Pont Polymers, Newark, DE, 1991. 9. Torlon, Engineering Polymers Design Manual, Amoco Performance Products, Atlanta, GA, 1993. 10. Ekonol PTFE Blends, Tech. Bull. Form C-1226, Kennecott Corporation, Sanborn, NY, 1981. 11. Sung, N. and Suh, N.P., Effect of fiber orientation on friction and wear of fiber reinforced polymeric composites, Wear, 53, 129, 1979. 12. Lancaster, J.K., The effect of carbon fibre reinforcement on the friction and wear of polymers, Br. J. Appl. Phys., 1, 549, 1968.

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13. Arkles, B., Theberge, J., and Schireson, M., Wear behavior of thermoplastic polymer-filled PTFE composites, Lub. Eng., 33, 33, 1977. 14. Hong, M. and Pyun, S., Effect of fluorinated ethylene propylene copolymer on the wear behavior of polytetrafluoroethylene, Wear, 143, 87, 1991. 15. Tanaka, K. and Kawakami, S., Effect of various fillers on the friction and wear of polytetrafluoroethylene-based composites, Wear, 79, 221, 1982. 16. Gong, D., Zhang, B., Xue, Q., and Wang, H., Investigation of adhesion wear of filled polytetrafluoroethylene by ESCA, AES and XRD, Wear, 137, 25, 1990. 17. Pooley, C.M. and Tabor, D., Friction and molecular structure: the behavior of some thermoplastics, Proc. R. Soc. London Ser. A, 329, 251, 1972. 18. Lancaster, J.K., Composite self-lubricating bearing materials, Proc. Inst. Mech. Eng., 182, 33, 1967–1968. 19. Theberge, J., Silicones improve self-lubricating thermoplastics, Mach. Des., Sept. 6, 1984, p. 108. 20. Cirino, M., Friedrich, K., and Pipes, R.B., The effect of fiber orientation on the abrasive wear behavior of polymer composite materials, Wear, 121, 127, 1988. 21. Mens, J.W.M. and de Gee, A.W.J., Friction and wear behavior of 18 polymers in contact with steel in environments of air and water, Wear, 149, 255, 1991. 22. Fusaro, R.L., Geometrical aspects of the tribological properties of graphite fiber reinforced polyimide composites, ASLE Trans., 26, 209, 1983. 23. Booser, E.R., Bearing materials, Kirk-Othmer Encyclopedia of Chemical Technology, Vol. 4, John Wiley & Sons, 4th ed., 1992. 24. Santner, E. and Czichos, H., Tribology of polymers, Tribol. Int., 22, 103, 1989. 25. Blanchet, T.A. and Kennedy, F.E., Sliding wear mechanism of polytetrafluoroethylene (PTFE) and PTFE composites, Wear, 153, 229, 1992. 26. Blanchet, T.A. and Kennedy, F.E., The development of transfer films in ultra-high molecular weight polyethylene/stainless steel oscillatory sliding, Tribol. Trans., 32, 371, 1989. 27. Gillespie, L.H., Saxton, D.O., and Chapman, F.M., New design data for FEP, TFE Part 2 — Thermal, wear and electrical properties, Mach. Des., Feb 18, 1960, p. 156. 28. Abarou, S., Play, D., and Kennedy, F.E., Wear transition of self-lubricating composites used in dry oscillating applications, ASLE Trans., 30, 269, 1987. 29. McLaren, K.G. and Tabor, D., Visco-elastic properties and the friction of solids, Nature, 197, 856, 1963. 30. Tanaka, K. and Nagai, T., Effect of counterface roughness on the friction and wear of polytetrafluoroethylene and polyethylene, Wear of Materials, ASME, New York, 1985, 397. 31. Hu, T. and Eiss, N.S., The effects of molecular weight and crystallinity on wear of polytetrafluoroethylene, Wear of Materials, ASME, New York, 1983, 636. 32. Tetrault, D.M. and Kennedy, F.E., Friction and wear of ultrahigh molecular weight polyethylene on Co-Cr and titanium alloys in dry and lubricated environments, Wear, 133, 295, 1989. 33. Fusaro, R.L., Friction, wear, transfer and wear surface morphology of ultrahigh-molecular-weight polyethylene, ASLE Trans., 28, 1, 1985. 34. Bowers, R.C., Clinton, W.C., and Zisman,W.A., Frictional properties of polymers, Mod. Plast., 31(6), 131, 1954. 35. O’Rourke, J.T., Fundamentals of friction, PV, and wear of fluorocarbon resins, Mod. Plast., 43(1), 161, 1965. 36. Kapoor, A. and Bahadur, S., Transfer film bonding and wear studies on CuS-Nylon composite sliding against steel, Tribol. Int., 27, 323, 1994. 37. Blanchet, T.A. and Han, S., unpublished data. 38. Durand, J.M., Vardavoulias, M., and Jeandin, M., Role of reinforcing ceramic particles in the wear behavior of polymer-based model composites, Wear, 181–183, 833, 1995. 39. Bohm, H., Betz, S., and Ball, A., The wear resistance of polymers, Tribol. Int., 23, 399, 1990. 40. Ludema, K.C. and Tabor, D., The friction and visco-elastic properties of polymeric solids, Wear, 9, 329, 1966. 41. Janczak, K.J., Janczak, T., and Slusarski, L., Friction and wear of polymer composite material, Wear, 130, 93, 1989. 42. Tanaka, K., Effects of various fillers on the friction and wear of PTFE-based composites, Friction and Wear of Polymer Composites, K. Friedrich, Ed., Elsevier Science Publishers, Amsterdam, 1986, 137.

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43. Axen, N. and Jacobson, S., A model for the abrasive wear resistance of multiphase materials, Wear, 174, 187, 1994. 44. Bahadur, S. and Gong, D., Formulation of the model for optimal proportion of filler in polymer for abrasive wear resistance, Wear of Materials, ASME, New York, 1991, 177. 45. 1989 guide to selecting engineering materials, ASM Metals Progress, ASM International, Metals Park, OH, 1989. 46. Jamison, W.E., Plastics and Plastic Matrix Composites, CRC Handbook of Lubrication and Tribology, Vol, 3, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1994, 121. 47. Whittington’s Dictionary of Plastics, Technomic Publishing, Lancaster, PA, 1993. 48. Rosato, D.V., Plastics Encyclopedia and Dictionary, Oxford University Press, London, 1993. 49. Encyclopedia of Polymer Science and Engineering, Mark, H.F., Bikales, N.M., Oberberger, C.G., Menges, G., and Kroschwitz, J.I., Eds., John Wiley & Sons, New York, 1987. 50. Handbook of Plastic Materials and Technology, Rubin, I.I., Ed., John Wiley & Sons, New York, 1990. 51. Handbook of Plastics, Elastomers and Composites, Harper, C.A., Ed., McGraw-Hill, New York, 1992. 52. Suh, N.P. Tribophysics, Prentice-Hall, Englewood Cliffs, NJ, 1986, 261.

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53 Properties of Advanced Ceramics S. Frank Murray CONTENTS Introduction.....................................................................................................................................................563 Physical Properties..........................................................................................................................................563 The Lubrication of Ceramics.......................................................................................................................566 Metal Alloy-Ceramic Combinations...........................................................................................................569 References.........................................................................................................................................................570

INTRODUCTION For tribological applications, ceramics offer many advantages over metals, including: high temperature strength, hot hardness, dimensional stability, and low density. Because of these attributes, there is considerable interest in the friction and wear of these materials for high-temperature bearing applications. During the past 30 years, improvements in the purity, manufacturing techniques, and finishing have resulted in more consistent and reliable products. However, the fact remains that ceramics are brittle materials with limited fracture toughness.1 Engineering data in the literature are sometimes questionable, ceramic compositions are not always well defined, and the environmental effects may not have been considered. Humidity and atmospheric contaminants have significant effects on sliding behavior.2–4 Better specifications, and test methods for material characterization, are needed. Ceramic products are formed by sintering and/or pressing fine powders of materials such as oxides,nitrides,carbides,or silicides.Uniform particle sizes,ranging from 1 to 3 µm (microns) for silicon nitride up to 40 to 60 µm for zirconia, are essential for uniform properties. Bonding and high density are usually achieved by the addition of oxide sintering aids to densify and fuse the hard particles. These sintering aids become the grain boundaries between the particles and determine the high temperature strength, which falls rapidly as these ceramic “glues” soften at high temperature. While ceramics are brittle below the softening range, progress is being made in improving toughness.

PHYSICAL PROPERTIES Table 1 lists typical properties of ceramic cutting tool inserts. The data were obtained from a number of sources including References 5 to 10. Actual numbers will depend on the composition, grain size, sintering conditions, etc. These tool inserts are used to cut superalloys, cast alloys, and hardened steels at very high velocities and depths of cut. Under those conditions, the temperatures generated at the cutting contacts are high enough to soften any metallic cutting tool alloy or cemented carbide matrix. Since the hot pressed (HP) sialon appears to be more widely Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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used as a cutting tool than silicon nitride, properties of the latter are shown in Table 3. The performance of these ceramic cutting tools has been reviewed by several authors,5 and they provide recommendations about the compatibility of the cutting tool with the workpiece alloy. While these materials are now being used primarily for metal removal, rather than bearing or seal applications, they are the most advanced ceramics for tribological applications and, in time, they will find uses in bearings and seals where they will be sliding or rolling against themselves or against metals. Table 2 lists properties of abrasive grit materials which are bonded to grinding wheels, abrasive belts, or abrasive paper. Various types of aluminum oxide, silicon carbide, diamond, and cubic boron nitride (CBN) are members of this class. A general discussion of their physical characteristics is presented in Reference 11. Solid cutting tools of diamond or CBN are also available, but are very costly. Layers of polycrystalline diamond or CBN can be bonded to the cutting edge to make a tool that can be resharpened. Table 3 is a list of ceramic materials being used extensively as bearings and seals. Of these, aluminum oxide (Al2O3) and silicon carbide (SiC) are well established. Silicon nitride (Si3N4) has been the subject of a considerable research effort during the past 20 years.12,13 It is a very effective ball material with outstanding fatigue life when run with steel races in oil-lubricated high-speed rolling element bearings. As ceramics go, this material has very good fracture toughness. Partially stabilized zirconia (ZrO2 + Y2O3 or MgO) has attracted attention for use in the lowheat rejection diesel engine because of its low thermal conductivity and good sliding wear characteristics. However, it was found to be sensitive to thermal shock, always a concern with ceramic

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components that must operate at high temperature. Glaeser,6 pp. 140–142, described the following relationship of thermally sensitive properties in the thermal shock resistance (TSR) of ceramics.

where σ κ E b

is the fracture stress is the thermal conductivity is the elastic modulus is the thermal expansion coefficient.

Thermal diffusivity is another significant parameter. This is a measure of the rate at which heat is dissipated from hot spots that would be generated when surfaces were sliding in contact at high speeds or stress levels:

where α κ ρ c

= thermal diffusivity, m2/s-1 = thermal conductivity, W/m ⋅ C = density, Kg/m3 = heat capacity, J/kg ⋅ C

Typical thermal data are compared in Table 4.

THE LUBRICATION OF CERAMICS This discussion is concerned with two classes of ceramics. One class includes the silicon ceramics such as silicon nitride (Si3N4), silicon carbide (SiC), and also the sialons, which are phases in the silicon-aluminum-oxygen-nitrogen (Si-Al-O-N) system. The second class covers the oxide ceramics such as aluminum oxide (Al2O3), zirconium oxide (ZrO2), and the partially stabilized zirconium Copyright © 1997 CRC Press, LLC.

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oxide (PSZ). There are a number of modifications which are being used to improve the characteristics ofAl2O3.Some involve the use ofhard particles such as titanium carbide (TiC) or reinforcing whiskers ofSiC to act as crack stoppers,8 p.525.Another variation uses the addition ofZrO2 (ZTA) to toughen the alumina by crystallographic techniques.14,15 There is still much to be learned about the differences between these classes of ceramics and their reactivities with the environment. Unlubricated ceramic combinations are being used successfully in some lightly loaded, lowspeed applications. However, it has generally been found that ceramics need some form of lubrication to prevent high friction and excessive wear.16–18 Except for very mild operating conditions, the wear rates for unlubricated ceramic couples (e.g., 1 × 10-4 to 1 × 10-6mm3/N · m) are orders of magnitude higher than the wear rates that could be tolerated in practical devices (e.g., 1 × 10-7 to 1 × l0-10 mm3/N · m, or less). Typical steady-state unlubricated coefficients of friction range from 0.5 to 1.0, although PSZ has given friction values as low as 0.25. Initially, low friction and smooth sliding take place until contaminating surface films are worn away. Then, the ceramic wear process begins with the microfracture of asperity contacts. As this damage progresses, loose wear debris becomes trapped in the contact area, resulting in three-body abrasion. In some cases, the debris is compacted and becomes bonded to one or both of the surfaces. Candidate lubricants for ceramic couples include: oils, greases, low shear strength solids, reactive vapors, and friction polymers. Adsorbed water molecules, liquid water, and aqueous solutions can also be effective lubricants in many applications.19–21 Table 5 is a list of the kinds of lubricants that have shown promise for use with ceramic bearings operating in various temperature ranges. Water-lubricated bearings and seals head the list.22 Water, either as a liquid or in the vapor state, can react with silicon nitride2,19 or aluminum oxide20 to form soft, protective surface films by a tribochemical process. Wear of this protective film improves the conformity and topography of the bearing surfaces.19,23,24 Note, however, that there are limiting loads and sliding speeds beyond which these films will break down. The wear rates will increase rapidly25,26 if the bearings do not generate hydrodynamic films to carry at least part of the load. Some evidence has been accumulated which indicates that the performance of other silicon ceramics, e.g., silicon carbide or the sialons, may also be affected by tribological reactions with water.3,23,24 This possibility needs further investigation. Zirconia (ZrO2) and PSZ are not recommended for use in water, or even in humid air, because of high wear rates,27–30 believed to be due to stress corrosion cracking.30 Both mineral and synthetic oils can provide boundary lubrication for some ceramic components. The oils and additives that lubricate steel surfaces appear to be just as effective with silicon nitride4,31 as with steel. On the other hand, oil-lubricated, self-mated zirconia (PSZ) sliding under boundary-lubricated conditions, may wear excessively because its low thermal conductivity inhibits the dissipation of frictional heat. Changing the combination to zirconia vs. hardened steel, or to some other high conductivity material, could resolve this problem of heat buildup.29 With Al2O3, it has been shown that polar compounds, such as fatty acids or soaps, can be adsorbed to form protective films.34 This does not seem to be the case with SiC. It has been hypothesized that polar compounds do not adsorb on SiC because the percentage of bonding in SiC is predominantly covalent (0.19 ionicity [Reference 8, p. 544]), while Al2O3 is ionic. This may be the correct explanation, but it does not explain why Si3N4, which also has a significant percentage of covalent bonding (0.3 ionicity [Reference 8, p. 531]), is so much more responsive to additive effects than either SiC or Al2O3. At temperatures above the 250 to 300°C level, the list of candidates has dwindled down to vapor phase lubrication,35,36 solid film lubricants,37–39 and oxide films formed by reaction with the environment.40–43 Development work is in progress on the first two approaches. The third technique, oxide film lubrication, will either require implantation of the lubricating components into the ceramic surfaces40,41 or the use of metal alloy/ceramic combinations that can react with the environment to form protective, lubricating oxides.

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METAL ALLOY-CERAMIC COMBINATIONS The feasibility of using a metal-ceramic bearing instead of all-ceramic combination seems to be an attractive solution, because it could reduce problems of brittle fracture as well as manufacturing costs. However, the metallic component must have sufficient strength and adequate corrosion resistance over the anticipated operating temperature range. It must also have the ability to form a low shear strength protective oxide film to prevent surface damage at high temperatures. Previous studies of metal-ceramic combinations42 have shown that transfer of the metal to the ceramic is the first step in the sliding process. Sliding was then metal vs. a work-hardened film of transferred metal. At room temperatures, coefficients of friction on the order of 0.6 to 0.7 were obtained. As the test temperatures were raised, both the original and the transferred metal surfaces were gradually oxidized. At high temperatures, depending on the type of oxide formed, the friction decreased to about 0.3 to 0.4, accompanied by smooth sliding. Figure 143 is a series of photographs showing the conversion from metal transfer to oxide film formation. The most promising alloys to use for beneficial oxide formation are nickel- and cobaltbase superalloys. Additional alloying elements, such as tungsten and molybdenum also contributed to performance by reacting to form soft, complex double oxides such as tungstates and molybdates. Confirming results have been reported44 using a nickel-base superalloy and comparing the performance of this alloy sliding vs. ceramics to the performance of the self-mated ceramic. Evaluations have also been made of Ni-Mo-Fe alloy sliding against Al2O3 in water.45

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FIGURE 1 Effect of temperature on lubrication by protective oxide film. Reciprocating pinon-flat test. Pin: Co-Mo-Cr-Si Laves phase alloy. Flat NBD 100 silicon nitride.

REFERENCES 1. Sims, C.T., Non-metallic materials for gas turbine engines, Advanced Materials and Processes, June 1991, 32–39. 2. Fischer, T.E. and Tomizawa, H., Interaction of tribochemistry and microfracture in the friction and wear of silicon nitride, Wear, 105, 29–45, 1985. 3. Ishigaki, H., Nagata, R., and Iwasa, M., Effects of adsorbed water on friction of hot pressed silicon nitride and silicon carbide at slow speed sliding, Wear, 121, 107–116, 1988. 4. Habeeb, J.J., Blahey, A.G., and Rogers, W.N., Wear and lubrication of ceramics, Intern. Conf. on Tribology-Friction, Lubrication and Wear, Vol. 1, Inst. Mech. Eng., London, 1st to 3rd, July 1987, pp. 555–564. 5. Ceramic Bull., Machining issue, 67(6), 991–1052, 1988. 6. Glaeser, W.A., Materials for Tribology, Tribology Series No. 20, Elsevier, New York, 1992. 7 Encyclopedia of Materials Science and Engineering, Vol. 1, Bever, M.B., Ed., MIT Press, Cambridge, MA, 1986.

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8. Komeya, K. and Matsui, M., High temperature engineering ceramics, chap. 10, in Structures and Properties of Ceramics, Swain, M, Ed., Materials Science and Technology, Vol. 11, Cahn, R.W., Haasen, P., and Kramer, E.J., Eds., Weinheim, New York, 1994. 9. CRC Practical Handbook of Material Science, Lynch, C.T., Ed., CRC Press, Boca Raton, FL, 1989. 10. DeVries, R.C., Cubic Boron Nitride: Handbook of Properties, G.E. Corporate R & D, No. 72CRD178, June 1972, 17 pages. 11. Gardinier, C.F., Physical Properties of Superabrasives, Ceramic Bull., 67(6), 1006–1009, 1988. 12. Bhushan, B. and Sibley, L.B., Silicon nitride rolling bearings for extreme operating conditions ASLE Trans., 25(4), 417–428, 1982. 13. Zaretsky, E.V., Ceramic bearings for use in gas turbine engines, ASME J. Eng. Gas Turbines Power, 111, 146–157, 1989. 14. Davidge, R.W., Perspectives for engineering ceramics in heat engines, High Temp. Technol., 5(1), 13–21, 1987. 15. He, C., Wang, Y.S., Wallace, J.S., and Hsu, S.M., Effect of microstructure on the wear transition of zirconia-toughened alumina, Wear, 162–164 Part A, 314–321, 1993. 16. Kato, K., Tribology of Ceramics, Wear, 136, 117–133, 1990. 17. Yust, C.S. and Carignan, F.J., Observations on the sliding wear of ceramics, ASLE Trans., 28(2), 245–252, 1985. 18. Sliney, H.E. and Dellacorte, C., The friction and wear of ceramic/ceramic and ceramic/metal combinations in sliding contact, Lubr. Eng., 50(7), 571–576, 1994. 19. Tomizawa, H. and Fischer, T.E., Friction and wear of silicon nitride and silicon carbide in water: Hydrodynamic lubrication at low sliding speed obtained by tribochemical wear, ASLE Trans., 30(1), 41–46, 1987. 20. Gates, R.S., Hsu, S.M., and Klaus, E.E., Tribological mechanism of alumina with water, Tribol. Trans., 32(3), 357–363, 1989. 21. Perez-Unzuetta, A.J., Beynon, J.H., and Gee, M.G., Effect of surrounding atmosphere on the wear of sintered alumina, Wear, 146, 179–196, 1991. 22. Johnson, R.L. and Schoenherr, K., Seal wear, in Wear Control Handbook, Peterson, M.B. and Winer, W., Eds., ASME, New York, 1980, 727–753. 23. Andersson, P., Water-lubricated pin-on-disc tests with ceramics, Wear, 154, 37–47, 1992. 24. Andersson, P., Nikkila, A.P., and Lintula, P., Wear characteristics of water-lubricated SiC journal bearings in intermittent motion, Wear, 179, 57–62, 1994. 25. Ravikiran, A. and Pramila Bal, B.N., Water-lubricated sliding of Al2O3 against steel, Wear, 171, 33–39, 1993. 26. Takadoum, J., Tribological behavior of alumina sliding on several kinds of materials, Wear, 170, 285–290, 1993. 27. Scott, H.G., Friction and wear of zirconia at very low sliding speeds, Int. Conf. on Wear of Materials, Ludema, K.C., Ed., ASME, New York, 1985, 8–12. 28. Sasaki, S., The effects of the surrounding atmosphere on the friction and wear of alumina, zirconia, silicon carbide and silicon nitride, Wear, 134, 185–200, 1989. 29. Zum Gahr, K.H., Sliding wear of ceramic-ceramic, ceramic-steel and steel-steel pairs in lubricated and unlubricated contact, Wear, 133, 1–22, 1989. 30. Fischer, T.E., Anderson, M.P., Jahanmir, S., and Salher, R., Friction and wear of tough and brittle zirconia in nitrogen, air, water, hexadecane, and hexadecane containing stearic acid, Wear, 124, 133–148, 1988. 31. Braza, J.F., Licht, R.H., and Tilley, E., Ceramic cam roller follower simulation tests and evaluation, Trib. Trans., 35(4), 595–602, 1992. 32. Dufrane, K.F., Wear performance of ceramics in ring/cylinder applications, J. Am. Ceram. Soc., 72(4), 691–695, 1989. 33. Hilton, M.R. and Fleisschauer, P.D., Lubricants for high-vacuum applications, in ASM Handbook, Vol. 18, Friction, Lubrication and Wear Technology, ASM International, Materials Park, OH, 1992, 150–161. 34. Studt, P., Influence of lubricating oil additives on friction of ceramics under conditions of boundary lubrication, Wear, 115, 185–191, 1987. 35. Hanyaloglu, B. and Graham, E.E., Vapor phase lubrication of ceramics, Lubr. Eng., 50(10), 814–820, 1994.

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36. Smith, J.C., Furey, M.J., and Kajdas, C., An exploratory study of vapor-phase lubrication of ceramics by monomers, Wear, 181–183, 581–593, 1995. 37. Murray, S.F. and Calabrese, S.J., Effect of solid lubricants on low speed sliding behavior of silicon nitride at temperatures to 800°C, Lubr. Eng., 49(12), 955–964, 1993. 38. Dellacorte, C. and Sliney, H.E., Composition optimization of self-lubricating chromium carbide based composite coatings for use to 760°C, ASLE Trans., 30(1), 77–83, 1987. 39. Vleck, B.L., Sargent, B.L., and Lauer, J.L., Lubrication of ceramic contacts by surface-deposited pyrolytic carbon, Lubr. Eng., 49(6), 463–471, 1990. 40. Lankford, J., Wei, W., and Kossowsky, R., Friction and wear behavior of ion beam modified ceramics, J. Mater. Sci., 22, 2069–2078, 1987. 41. Gangopadhyay, A.K., Fine, M.E., and Cheng, H.S., Friction and wear characteristics of titanium and chromium doped polycrystalline alumina, Lubr. Eng., 44(4), 330–334, 1988. 42. Peterson, M.B. and Murray, S.F., Frictional behavior of ceramic materials, Metals Eng. Q., 7(2), 22–29, 1967. 43. Murray, S.F. and Calabrese, S.J., Low speed sliding behavior of metal-ceramic couples at temperatures up to 800°C, Lubr. Eng., 49(5), 387–397, 1993. 44. Sliney, H.E. and Dellacorte, C, The friction and wear of ceramic/ceramic and ceramic/metal combinations in sliding contact, Lubr. Eng., 50(5), 571–576, 1994. 45. Holzhauer, W., Johnson, R.L., and Murray, S.F., Wear characteristics of water-lubricated Al2O3metal sliding couples, 1981 Wear of Materials Conference, Kudema, K.C. Ed., ASME, New York, 1981, 676–684.

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and Wear Data on Advanced 54 Friction Ceramics Amp Gangopadhyay CONTENTS α-Alumina.......................................................................................................................................573 Alumina...........................................................................................................................................574 Silicon Nitride.................................................................................................................................574 Silicon Carbide................................................................................................................................575 Partially Stabilized Zirconia (PSZ)..............................................................................................575 Aluminum Titanate........................................................................................................................576 Al2O3-TiC........................................................................................................................................576 Al2O3-20 vol% SiC.........................................................................................................................577

α-ALUMINA TEST TYPE: PIN ON DISK, NONLUBRICATED Test Conditions: Sliding in air, 50 + 5% relative humidity, at 0.2 m/s sliding velocity for a total sliding distance of 250 m under 10-N load Material Definitions: Ball: Sintered α-alumina, 9.5–11.1 mm diameter, 1650 HV hardness Disk: 100Cr6 steel, polished Ra 0.06–0.1 mm Test Results: Friction coefficient, 0.68 Wear rate of ball, 9 × 10-8 mm3/m Wear rate of disk, 8.33 × 10-6 mm3/m Reference: Anderson, P. and Holmberg, K., Limitations on the use of ceramics in unlubricated sliding applications due to transfer layer formation, Wear, 175, 1–8, 1994. TEST TYPE: BALL ON FLAT, NONLUBRICATED Test Conditions: Reciprocating sliding in air, 30–50% relative humidity, at 1.4 mm/s sliding, for a total sliding distance of 10 m under 10–100N load at 23–1000°C Material Definitions: Ball: α-Alumina (99.5%); polished, roughness 0.01 mm RMS; 14.7 GPa hardness Disk: α-Alumina (99.8%); polished, roughness 0.1 mm RMS; 15.0 GPa hardness Test Results: Stroke length, 10 mm Mass loss of ball, 0.02–0.24 mm3 Friction coefficient, 0.35–0.9

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Mass loss of disk, 0.01–0.35 mm3 Reference: Dong, X., Jahanmir, S., and Hsu, S.M., Tribological characteristics of α-alumina at elevated temperatures, J. Am. Ceram. Soc., 74(5), 1036–1044, 1991. TEST TYPE: BLOCK ON RING, LUBRICATED WITH MINERAL OIL WITHOUT ADDITIVES Test Conditions: Steady sliding in air at 1.66 m/s sliding velocity for a total sliding distance of 106 km under 197-N load Material Definitions: Block: 19 mm × 9.5 mm × 3 mm α-alumina; 2,150 kg/mm2 hardness Ring: 62.5 mm o.d. × 50.8 mm i.d.; 52100 steel; Rc 62 Test Results: Wear rate of block, 2.5 × 10-8 mm3/m ⋅ N Friction coefficient, 0.12 Reference: Gangopadhyay, A.K., Fine, M.E., and Cheng, H.S., Friction and wear characteristics of titanium and chromium doped polycrystalline alumina, Lub. Eng., 44(4), 330–334, 1988.

ALUMINA TEST TYPE: THRUST WASHER, LUBRICATED WITH TAP WATER Test Conditions: Steady sliding in air at 0.5 m/s sliding velosity under 290 Kpa contact stress for 48 h at room temperature Material Definitions: Rotating washer: alumina (85%) with 40 µin. Ra finish Lower washer: 440 C steel; 58–60 Rc with 3–4 µin. Ra finish Washer size: 31.75 mm o.d., 12.7 mm i.d., 7.62 mm thick Test Results: Wear track diameter, 11.7 mm Steady-state wear rate of 440 C steel washer, 0.5 mg/h Friction coefficient, 0.05–0.2 Reference: Holzhauer, W., Johnson, R.L., and Murray, S.F., Wear characteristics of water lubricated Al2O3-metal sliding couples, Wear of Materials, San Francisco, Ludema, K.C., Ed., 1981, 676–684.

SILICON NITRIDE TEST TYPE: PIN ON DISK, NONLUBRICATED Test Conditions: Steady sliding in dry air at 1 mm/s sliding velocity for a total sliding distance of 3 m under 0.5–30-N load Material Definitions: Pin: 3-mm tip radius hot-pressed Si3N4 Disk: Polished to 5 nm rms hot-pressed Si3N4 Test Results: Wear rate of pin, 3 × 10-11 kg/N ⋅ m Friction coefficient, 0.8 ± 0.05 Wear rate of disk, 3.3 × 10-9 kg/N ⋅ m Reference: Fischer, T.E. and Tomizawa, H., Interaction of tribochemistry and microfracture in the friction and wear of silicon nitride, Wear, 105, 29–45, 1985.

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TEST TYPE: ROLLER ON RING LUBRICATED WITH MINERAL OIL Test Conditions: Steady sliding in air at 3.12 m/s sliding velocity for a total sliding distance of 110 km under 1112-N load Material Definitions: Roller: 17.8 mm o.d. × 7.5 mm i.d., 12.7 mm long hot-pressed silicon nitride; 17.9 GPa hardness, grain size, 0.7 µm Ring: 30 mm o.d., 30-mm long nodular cast iron; Rc 56 Test Results: Friction coefficient, 0.046 Wear rate of roller, 5 × 10-10 mm3/N ⋅ m Reference: Braza, J.F., Cheng, H.S., and Fine, M.E., Silicon nitride wear mechanisms: rolling and sliding contact, Tribol. Trans., 32(4), 439–446, 1989.

SILICON CARBIDE TEST TYPE: PIN ON FLAT, NONLUBRICATED Test Conditions: Reciprocating sliding in air at 1 mm/s sliding velocity under 10–100N load Material Definitions: Ball: 10 mm dia 52100 steel; 8.2 GPa hardness Disk: Polished SiC, Ra 0.1 µm finish; 24 GPa hardness Test Results: Stroke length, 4 mm Wear scar dia. of ball, 300–600 µm Friction coefficient, 0.4–0.6 Reference: Kapsa, P., Maurin-Perrier, P., and Pijard, B., Frictional properties of silicon carbide against steel and sapphire in dry conditions, Proc. Inst. Mech. Eng. Int. Conf. Tribology, London, July 1 to 3, Vol. 2, 1987, 687–694.

PARTIALLY STABILIZED ZIRCONIA (PSZ) TEST TYPE: PIN ON RING LUBRICATED WITH MINERAL OIL WITHOUT ADDITIVES Test Conditions: Steady sliding in air, 42–56% relative humidity at 0.77 m/s sliding velocity for a total sliding distance of 8.7 km under 200-N load Material Definitions: Pin: 5 × 5 mm2 Mg-PSZ (96% ZrO2); ground, Ra 0.3 µm; 748 HV hardness Ring: Hardened steel; ground, Ra 0.25–0.36 µm; 470 HV hardness Test Results: Wear rate of pin, 2 × 105 µm3/km Friction coefficient, 0.12 Wear rate of ring, 3 × 106 µm3/km Reference: Zum-Gahr, K.H., Sliding wear of ceramic/ceramic, ceramic/steel and steel/steel pairs in lubricated and unlubricated contact, Wear of Materials, Denver, Ludema, K.C., Ed., 1989, 431–440. TEST TYPE: PIN ON RING LUBRICATED WITH MINERAL OIL WITHOUT ADDITIVES Test Conditions: Steady sliding in air, 42–56% relative humidity at 0.77 m/s sliding velocity, for a total sliding distance of 8.7 km under 200–N load Material Definitions: Pin: 5 × 5 mm2 Mg-PSZ (96% ZrO2); ground, Ra 0.3 µm; 748 HV hardness Copyright © 1997 CRC Press, LLC.

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Ring: Mg-PSZ (96% ZrO2); ground, Ra 0.3 µm; 748 HV hardness Test Results: Wear rate of pin, 1 × 108 µm3/km Friction coefficient, 0.15 Wear rate of ring, 7 × 108 µm3/km Reference: Zum-Gahr, K.H., Sliding wear of ceramic/ceramic, ceramic/steel and steel/steel pairs in lubricated and unlubricated contact, Wear of Materials, Denver, Ludema, K.C., Ed., 1989, 431–440.

ALUMINUM TITANATE TEST TYPE: PIN ON RING LUBRICATED WITH MINERAL OIL WITHOUT ADDITIVES Test Conditions: Steady sliding in air, 42–56% relative humidity at 0.77 m/s sliding velocity for a total sliding distance of 8.7 km under 200–N load Material Definitions: Pin: 5 × 5 mm2 pressed Al2TiO5; ground, Ra 0.3 µm; 200 HV hardness Ring: Hardened steel; ground, Ra 0.25–0.36 µm; 470 HV hardness Test Results: Wear rate of pin, 8 × 108 µm3/km Friction coefficient, 0.17 Wear rate of ring, 8 × 108 µm3/km Reference: Zum-Gahr, K.H., Sliding wear of ceramic/ceramic, ceramic/steel and steel/steel pairs in lubricated and unlubricated contact, Wear of Materials, Denver, Ludema, K.C. Ed., 1989, 431–440.

AL2O3-TIC TEST TYPE: BLOCK ON RING, NONLUBRICATED Test Conditions: Steady sliding in air at 2.8–5.4 m/s sliding velocity for a total sliding distance of 2320 m under 267–N load Material Definitions: Block: Al2O3-TiC Ring: 4340 steel Test Results: Wear rate of block, (1.2–1.8) × 10-5 mm3/m Reference: Mehrotra, P.K., Mechanisms of wear in ceramic materials, Wear of Materials, Virginia, Ludema, K.C., Ed., 1983, 194–201. TEST TYPE: ROLLER ON RING LUBRICATED WITH MINERAL OIL Test Conditions: Rolling in air at 3.12 m/s sliding velocity for a total sliding distance of 1150 km under 1112–N load Material Definitions: Roller: A12O3 with 45 wt % TiC; 17.8 mm o.d. × 7.5 mm i.d., 12.7 mm long; grain size 1–2 µm, 18.1 GPa hardness Ring: Nodular cast iron; 30 mm o.d., 30 mm long; Rc 56 Test Results: Friction coefficient, 0.005 Wear rate of roller, -5 × 10-11 mm3/N ⋅ m

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Reference: Gangopadhyay, A.K., Cheng, H.S., Braza, J.F., Harman, S., and Corwin, J.M., Tribological performance of ceramic cam followers, Friction and Wear of Ceramics, Jahanmir, S., Ed., Marcel Dekker, New York, 1994, 329–356.

AL2O3–20 VOL% SIC TEST TYPE: PIN ON DISK, NONLUBRICATED Test Conditions: Steady sliding at 400°C in air, 35–55% relative humidity at 0.1–0.5 m/s sliding velocity for a total sliding distance of 1–2 km under 2.2–8.9N load Material Definitions: Pin: Al2O3–20 vol% SiC; polished, Ra 0.1 µm; grain size ≤ 4 µm Disk: Al2O3–20 vol% SiC; polished, Ra 0.1 µm; grain size ≤ 4 µm Test Results: Wear track diameter, 18–30 mm dia. Wear rate of pin, 10-9–10-7 mm3/N ⋅ m Friction coefficient, 0.4–0.7 Reference: Yust, C.S. and Allard, L.F., Wear characteristics of an alumina-silicon carbide whisker composite at temperatures to 800°C in air, Tribol Trans., 32(3), 331–338, 1989.

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VII Tribological Surface Modifications

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Surface Treatments and 55 Tribological Coatings Francis E. Kennedy and Ursula J. Gibson CONTENTS Introduction.....................................................................................................................................................581 Thermal Transformation Hardening Processes.......................................................................................581 Thermal Diffusion Processes.......................................................................................................................584 Electroplating and Anodizing Processes....................................................................................................585 Thermal Spray Processes...............................................................................................................................586 Weld Hardfacing Processes...........................................................................................................................587 Vapor Deposition Processes.........................................................................................................................588 Ion Implantation Processes..........................................................................................................................591 References.........................................................................................................................................................592

INTRODUCTION One of the primary remedies for friction and wear problems is to treat the surface, usually to make it harder. A wide variety of tribological surface treatments is in use today, ranging from flame hardening and weld hardfacing processes, which have been in use for decades, to ion beam processes and new surface coating methods which have been developed in the past few years. Some surface treatment processes produce a change in the hardness of the existing surface, whereas in other cases a finite thickness of a coating material is added over the original surface as shown schematically in Figure 1. Choice between available surface treatments must also be based on the effect on the underlying bulk material. Of particular importance is whether the temperature reached would have a deleterious effect on the mechanical properties of the substrate material. A list of the available processes for treating steel surfaces is given in Table 1, along with the temperature reached during treatment and the characteristics of the treated surfaces. Although most of the hardening processes listed are suitable only for ferrous surfaces, the coating processes can be used for other substrates as well. Each of the processes is covered in the following discussion and further details may be found in References 1 through 14.

THERMAL TRANSFORMATION HARDENING PROCESSES The thermal transformation hardening processes summarized in Table 2 are the simplest and most common ways to harden ferrous surfaces without adversely affecting the properties of the bulk of the material. The processes involve heating the surface rapidly, transforming it to austenite, and then quenching it to form martensite. The source of heat can be one of the following: • An oxyacetylene or oxypropane flame impinging on the surface (flame hardening) Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 A schematic comparison of surface treatment processes.

• Eddy currents induced by a high-frequency electric field from a coil around the part (induction hardening) • A beam from a high-power (1 to 15 kW) CO2 laser (laser hardening) • A focused electron beam (electron beam hardening). The latter two techniques are characterized by rapid localized heating from heat sources of high energy density and by rapid self-quenching as soon as the heat source moves away. The heat sources in flame and induction hardening cover a larger area and have less energy density; a separate quenching process is required to produce martensite in the treated layer. The depth and uniformity of the hard layer depend on the rate and method of heating and quenching, while the resulting Copyright © 1997 CRC Press, LLC.

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surface hardness depends on the carbon content of the material. Materials suitable for thermal transformation hardening include medium carbon steels, low or medium alloy (C-Mn, Cr-Mo, and Ni-Cr-Mo) steels, gray cast iron, pearlitic malleable cast iron, and nodular cast iron. A basic requirement of thermal transformation hardening is that the substrate be capable of martensite formation upon quenching. This requires an adequate carbon content, as is shown in Figure 2. The alloying additions to low alloy steels enable a slight increase in hardness, up to 100 HV above the values shown in Figure 2 for a given carbon content.

FIGURE 2 Hardness of untempered martensite as function of carbon content for plain carbon steels.1 High alloy steels are usually not hardened by thermal transformation processes because of the slow rate at which they austenitize and their tendency to suffer from quench cracks. Plain carbon steels with carbon contents above 0.6% also are susceptible to quench cracking, particularly with Copyright © 1997 CRC Press, LLC.

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flame or induction heating processes, owing to the low toughness of the hardened layer. Carbon steels with low carbon content (less than 0.3%) are not too susceptible to quench cracking, but the rate at which they must be cooled to produce martensite is so high that it cannot be easily attained by thermal transformation processes, particularly flame or induction hardening. Each of the thermal transformation processes is characterized by a short process time, and all except electron beam hardening (which requires a moderate vacuum) can be done in air. A tempering of the hardened surface is recommended after each of the thermal transformation processes.

THERMAL DIFFUSION PROCESSES These processes involve the diffusion of atoms into surfaces to create a hard layer. (See Table 3.) In the most widely-used, carburizing (or case hardening), carbon diffuses into a low-carbon steel surface to produce a hard, carbon-rich “case”. The source of carbon can be either a hydrocarbon or carbon monoxide gas in air (gas carburizing) or vacuum (vacuum carburizing), a salt bath with carbon-containing salts such as cyanide (salt bath carburizing), a packed bed of coke or charcoal (pack carburizing), or a glow discharge of methane plasma (plasma carburizing).2 T he hardness and thickness of the case depend on the temperature, exposure time, and source of carbon. Carburizing is most effective with low carbon steels (0.10 to 0.25% carbon), and the result is enhanced carbon concentration of between 0.7 and 0.9% in the surface layer, with a surface hardness of between 800 and 900 HV (see Figure 2). Carbonitriding is a process similar to carburizing which involves the simultaneous diffusion of carbon and nitrogen atoms into the surface of low carbon (5 µm, is usually indicative of severe wear. An unusually high level of dissolved metals, such as Mg, is also indicative of severe corrosion problems. Based on historical data specific for each engine in service, wear metal guidelines can Copyright © 1997 CRC Press, LLC.

Wear Monitoring and Metal Analysis

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be established for determining engine health. Rapid wear metal production even at low concentrations is an indication of impending oil wetted component failure. Table 2 lists critical wear metals analyzed for by SOAT and their possible metallurgical source. An example of Fe or Ti wear metal guidelines for the F-100 gas turbine engine is as follows: • Normal SOAT values, 15 If the engine is operating normally, the wear metal concentrations should be less than 12 ppm and the concentration increase 20 µm, >30 µm, >40 µm, and >50 µm. The number categories and the sizes are specific to the instrument employed and its calibration. An International Standards Organization (ISO) standard has gained wide acceptance for reporting particle count information in the form x/y, where x is the number >5 µm/ml and y is the number >15 µm/p/ml using a special scale. The scale uses numbers from 1 to 24 as a shorthand to describe the number per milliliter, with 24 being the worst case (80,000 to 160,000 particles per ml) and 1 being extremely clean (from 0.01 to 0.02 particles per ml). Particle counters typically use one of two different techniques for particle counting in lubricant and hydraulic systems, light-blocking and light-scattering. Light-scattering sensors are ideal for counting and sizing light-colored particles since they reflect or scatter light. They operate by detecting the amount of light reflected or scattered by a particle. As a liquid sample enters the sensor, it passes through the “view volume,” an area of intense laser light (Figure 6). Particles in the sample scatter bursts of light into a series of collection optics that focus the burst of light onto a solid-state photo diode. The photo diode converts the light to electrical pulses where the quantity and height correspond to quantity and size of particles.9

FIGURE 6 Light-scattering laser sensor. Light-blocking sensors operate by detecting the “shadow” created by a particle (Figure 7). As a liquid sample passes through the “view volume,” an area of intense laser light, particles in the sample momentarily block the laser light. A solid-state photo diode detects the momentary decrease in light and creates a corresponding electrical pulse that is proportional to particle size.9 While both light-scattering and light-blocking sensors can be used for counting particles, lightblocking sensors are preferred because they have simpler optics resulting in lower cost, are more effective at counting dark particles, remain calibrated over a larger range of flow rates, handle a larger concentration of wear and contaminate particles, and have longer service life. On the other hand, instruments with the light-scattering sensor can detect smaller particles.

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FIGURE 7 Light-blocking laser sensor. ON-SITE DEBRIS MONITORS Several other relatively simple instruments for wear particle analysis are for fast on-site screening of used oil samples or for selecting those samples requiring more detailed laboratory analysis. Since they are based on magnetism and permeability, their limited sensitivity restricts their application to systems with high wear modes and high particle concentrations such as gear boxes and transmissions. These instruments do not involve laboratory techniques and therefore will not be described in detail.

LUBRICANT CONDITION MONITORING The second part of an effective oil analysis program is lubricant condition monitoring to determine the effectiveness and remaining life of the lubricant based on degradation and contamination analysis. The number and type of tests performed on a used oil sample vary with the type of oil and the type of machine being monitored. Physical property tests performed by the typical used oil analysis laboratory, often using modifications of ASTM procedures to reduce analysis time, include the following: viscosity, TBN (total base number), TAN (total acid number), water content (Karl Fischer), fuel dilution, and insolubles. VISCOSITY In the laboratory, ASTM D 445,10 is the most commonly employed test. This is done using capillary tubes suspended in precisely controlled temperature baths, at either 40°C, or 100°C, or often at both temperatures, depending on the application. Oil is allowed to flow through a capillary tube and time of passage between two calibration marks on the capillary tube is measured. Sample viscosity in centistoke (cSt) units is a multiple of the time measurement and the capillary tube calibration factor. Analysis time is a function of oil viscosity and can vary from one to several minutes. Most laboratory viscometer baths can accommodate from 4 to 8 capillary tubes and throughput is increased by analyzing multiple samples at the same time. Semiautomatic and completely automatic viscometers that meet ASTM D 445 are also available for used oil analysis. Typical semiautomatic viscometers require one manual function such as the insertion of the sample in the capillary tube, or cleaning of the capillary tubes. Fully

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automatic viscometers use sample changers and are completely automated such that unattended operation is possible. The high capital costs are justified with a large sample load because of reduced labor costs. Viscosity may also be measured on site by the electrical power required to maintain constant amplitude in an oscillation viscometer. The test is commonly performed at ambient temperature, which is measured with a thermocouple built into the sample probe. Using density and viscosity index data, viscometer readings can be converted by computer to centistokes at the standard reference temperatures of 40°C and 100°C. A temperature bath can also be used to directly measure viscosity at these two temperatures. This instrument is not as accurate as a capillary tube in a controlled temperature bath, but accuracy is sufficient for condition monitoring applications. The advantage is its ease of use, speed of measurement (less than 1 minute per sample), and ruggedness of design. Disadvantages are that it is fairly expensive, does not meet the requirements of ASTM D 445, and is not as accurate as capillary tube viscometers. NEUTRALIZATION NUMBER Neutralization numbers are reported as either TAN or TBN, depending on the lubricant or the application. TAN is a measure of the total amount of acid products present in the lubricant. TAN values of new oils tend to be less than 1 and gradually rise with oxidation of the lubricant due to operation at high temperatures. A high acid number is an indication of oil breakdown. ASTM D 974 determines the acidity level of the lubricant by observing a color change as an indicator solution of potassium hydroxide (KOH) is added. TAN is expressed as milligrams of KOH required to neutralize a gram of oil (mgKOH/g). The method is accurate to 15%.11 ASTM 664, used with oils that are opaque and too dark to use a color change indicator, determines lubricant acidity by measuring the potentiometric change as KOH is added. TBN, the opposite of TAN, measures the alkalinity reserve remaining in the lubricant — in other words, its ability to neutralize corrosive acids that are formed during operation. TBN procedure ASTM D 2896 measures alkalinity and results are also expressed in mgKOH/g. Several manufacturers market titrators capable of performing automatic TAN and TBN titrations. The instruments are computer controlled and small enough for counter top operation. Automatic samplers are available since some titrations require a fair amount of time. Both TAN and TBN require the use of organic solvents which are not only expensive to purchase, but also expensive to dispose of properly. WATER CONTENT The amount of water suspended in a used oil sample can be assessed through tests that vary in their detection ability. A visual test will show perhaps 500 ppm or greater concentration by the presence of turbidity or cloudiness.7 Laboratory analyses of water vary from the simple “crackle” test to the very accurate coulometric Karl Fischer titration. The “crackle” test is performed by placing a few drops of the used oil sample on a hot surface (about 300°F). Any water present will quickly vaporize with a crackling or popping sound. The test is subjective and does not quantify the amount of water present. The three ASTM methods used for the analysis of water in used oil are ASTM D 95, ASTM D 1744, and ASTM D 4928. ASTM D 95, which determines water in oil by distillation, is time consuming and not practical for most used oil analysis applications. It measures only high concentrations of water over 1000 ppm. ASTM D 1744 determines water content by reacting the oil sample volumetrically with Karl Fisher reagent to an electrometric end point. While the procedure can be carried out manually with the appropriate glassware, commercial instruments are automatic and can be purchased as stand alone products or as an ancillary device to titrators used for TAN and TBN.

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The preferred method to measure trace water levels in oil down to levels below 10 ppm is by coulometric Karl Fischer titration, ASTM D 4928. A number of instrument manufacturers offer automatic coulometric titrators which employ a potentiometric end point and generate Karl Fischer reagent electrolytically.12 The principal advantages of coulometric titrations are low detection limit and the elimination of problems associated with the preparation, standardization, and storage of standard solutions. For most used oil analysis applications, the volumetric method is adequate, with the coulometric preferred for production quality control. Neither method can be used with automatic sample changers because the decanted sample waiting for analysis will absorb atmospheric water giving spurious results. INSOLUBLES Among solid content tests, the only one that is easy to perform is the blotter test used to evaluate the concentration of insolubles and to estimate the dispersive power of the used oil.13 After one drop of used oil is placed on a special type of filter paper, visual interpretation of the spot left determines whether or not the oil is still dispersant in character or whether insoluble material has started to drop out of the oil. Although some automatic techniques to read the spot test have been developed, the method is still highly subjective and interpretation is something of an art. It applies to insolubles ranging from approximately 0.2 to 3.5% by weight. ASTM D 893 describes the centrifuge method to determine insolubles in used oil. Two procedures are available. In one, the used oil is mixed with pentane and then centrifuged. In the other, the used oil sample is mixed with pentane-coagulant solution and then centrifuged. In both procedures, the precipitate is washed, dried, and weighed to give the insolubles content. The centrifugation method became self-defeating when the modern dispersant additives in lubricants necessitated adding coagulants to the sample in order to effect a more complete separation of solids during centrifugation. The filtration method to measure insolubles starts with a weighed amount of used oil sample which is diluted with pentane or heptane prior to being filtered through a membrane. Percentage of insolubles is calculated from the increased weight of the membrane after filtration and drying. While membrane filters can be selected according to pore size to measure insolubles above a preselected critical size, large particles may block the filter pores and thus introduce an error by holding back insolubles smaller than the pore size. This procedure uses highly flammable solvents. Thermogravimetry can also be applied for determination of insolubles in used oils. Automatic instruments are used for the test in which a weighed sample of used oil is heated in a stream of nitrogen in increments up to 650°C. After a few minutes, the sample weight is recorded and 10% of air is introduced into the stream of nitrogen to oxidize the carbon and soot. The air content is gradually increased to 100%. When the sample weight has stabilized, the test is complete. The primary advantage of the thermogravimetric test is the ability to quantify separately the combustion-formed carbon or soot and the noncombustible portions of the used oil sample. Since this technique can take as long as one hour per sample, thermogravimetric insoluble determination is rarely applied in a routine used oil analysis program. It is applied more often in support of lubricant formulation or engine design and testing programs. FOURIER TRANSFORM-INFRARED SPECTROSCOPY (FT-IR) This technique is starting to receive wide acceptance as a rapid method which can measure and quantify in less than one minute, the information shown in Table 9. A major breakthrough came with FTIR spectrometers capable of collecting an entire spectrum in half a minute or less and the tedious measuring of the spectrum by hand was replaced with attached computers. FT-IR spectrometers measure energy in the infrared region of the spectrum where individual types of chemical bonds absorb light of specific wavelengths. The basic FT-IR spectrometer has three components: a source, a Michelson interferometer, and a detector, Figure 8. The interferometer

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FIGURE 8 Schematic diagram of an FT-IR spectrometer. consists of a beamsplitter, a fixed mirror, and a moving mirror.14 In routine operation, broad band infrared light from the source is directed onto the beamsplitter where it is split into two beams of approximately equal energy. One beam is reflected from the fixed mirror and the other beam is reflected from the moving mirror. The reflected beams recombine at the beamsplitter. Here constructive or destructive interference takes place, depending on the position of the moving mirror relative to the fixed mirror. The net effect of the interferometer is to imbed an interference pattern into the infrared light which is next directed to the sample which selectively absorbs those wavelengths corresponding to the chemical bonds in the sample. Intensity of the infrared light passing through the sample is measured by the polychromatic infrared detector. The computer in the FT-IR spectrometer then performs a Fourier transform calculation to convert the time/intensity points to frequency/intensity points from which software extracts and reports information on lubricant degradation and contamination levels.

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Two approaches to FT-IR are in use for oil analysis. One applies subtraction prior to data extraction by comparing the used oil spectrum to the spectrum of the identical new oil.15 The difference between certain peaks of the two spectra is indicative of the change in the lubricant and is used for trending purposes. This technique necessitates either analysis of the new oil prior to the used oil, or recall of the new oil spectrum stored in the computer memory. The other technique is based solely on extracting the area under certain peaks of the spectrum. This is a trending technique based on changes in used oil samples over time. For FT-IR to be effective, the base lubricant for the used oil sample must be known for the subtraction approach, and the trend is defeated for either approach if the machine’s oil is changed. Primary advantage of FT-IR is speed of analysis and the number of properties reported in each analysis. Although often backed up with standard ASTM methods when more precision is required, it is excellent for trending techniques where changes in data are indicative of lubricant contamination and/or degradation. Used oil analysis using the FT-IR technique has been widely accepted over the past few years, and most instrument manufacturers provide custom software with analytical methods for the typical physical property analysis. Table 10 summarizes typical physical properties and their absorbance spectra.16,17

TURNKEY USED OIL ANALYSIS LABORATORIES Configuration and instrumentation of a laboratory will vary based on the machines being monitored and the sample work load. A full-service laboratory is shown in Figure 9, but the basic minimum components consist of an emission spectrometer, a Fourier transform-infrared spectrometer (FT-IR), and a viscometer. Each instrument sends its results to a data-based laboratory information management system for data storage, evaluation, and reporting. A rotating disc emission (RDE) spectrometer is the basic instrument recommended for routine measurement of the elemental concentration of wear metals, contaminants, and additives. It provides simplicity of operation, sensitivity to larger particles, freedom from diluting samples, and requires no gas or cooling water while completing analysis of approximately 20 elements in less than a minute. An atomic absorption spectrometer (AAS) is seldom used unless the sample volume is extremely low and cost per sample is not a consideration. An inductively coupled plasma (ICP) spectrometer is recommended only where absolute accuracy of results is important, such as quantification of additive elements in a lubricating-blending plant.

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FIGURE 9 Full-service turnkey used oil analysis laboratory. FT-IR spectrometers for used oil analysis have dedicated programs which extract lubricant degradation and contamination parameters from the measured spectrum of the used oil sample. The technique is fast, less than a minute per sample, and provides data on oxidation, nitration, sulfation, soot, fuel dilution, water and glycol contamination, and in some cases, additive depletion. As a fast trending technique, it has become a standard instrument in many high sample volume used oil analysis laboratories. A viscometer is the third required instrument in the basic turnkey used oil analysis system. Viscosity is the single most important physical characteristic of a lubricant since it determines load carrying ability as well as flow and heat flow characteristics. Manual viscometers are inexpensive and work well in low sample volume requirements. Automatic viscometers are readily available for various degrees of automation and unattended operation. In the basic system, measurements from each analytical instrument are sent to a central computer file where the results are incorporated into a history file for each unit (specific machine or sampling point on a machine). When tests are complete, the computer calls up the file of each unit and compares the results to a criteria matrix with allowable limits and to past analyses. In an automatic evaluation mode, records for samples with all data within limits are passed directly to the history file and a report with no recommended action is sent to the maintenance personnel. Samples with “out of limit readings” are flagged for review by the laboratory expert, who can then send a report with a maintenance recommendation to the maintenance personnel via telephone, telefax, or printed copy.

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This basic used oil analysis laboratory can be expanded as the analytical requirements of the laboratory or the sample work load increase. Ferrography, which magnetically separates the wear particles in an oil sample and arranges them according to size on a microscope substrate, gives important supplemental information on ferrous particles too large to be measured by routine spectrometric methods. Total acid number (TAN), total base number (TBN), and Karl Fischer water determination are three frequently performed ASTM tests for oil degradation and contamination. An automatic titrator is sometimes supplied with a turnkey system if more definitive information than that supplied by the FT-IR spectrometer is required. Particle count measurement is sometimes recommended, primarily for use with hydraulic systems or other clean lubricating oil systems such as those for turbines and compressors. With this added equipment, the used oil analysis laboratory combines the analytical speed required for large sample volumes with the additional capabilities of providing specialized ASTM based tests. It contains instruments and operating software designed specifically for used oil analysis with turnaround times of 24 to 48 hours to provide data trends used for effective machine condition monitoring. With expanding needs, a local area network (LAN) can be used to share information and additional tests can be added to match specific machinery monitoring needs.

REFERENCES 1. Lukas, A. and Anderson, D.P., Machine and lubricant condition monitoring for extended equipment lifetimes and predictive maintenance, paper presented at Int. Symp. Mining, September 15 to 17, 1992, Pretoria, South Africa. 2. Lukas, M., Lubricating oil analysis, Sawyer’s Turbomachinery Maintenance Handbook, Sawyer, J.W., Ed., Turbomachinery International Publications, Norwalk, CT, 1980, 8-1 to 8-20. 3. Sieber, J.R. and Salmon, S.G., Elemental analysis of lubricating oils and greases, Lubrication, Vol. 80, No. 1, 1994. 4. Lukas, M., Comparison of spectrometric techniques for the analysis of liquid gas turbine fuels, Trans. ASME, Vol. 115, 620-627, 1993. 5. Rhine, W.E., Saba, C.S., and Kaufmann, R.E., Metal particle detection capabilities of rotating disc emission spectrometers, Lubr. Eng., 42(12), 755, 1986. 6. Bruno, T. J. and Svorornos, P.D.N., CRC Handbook of Basic Tables for Chemical Analysis, CRC Press, Boca Raton, FL, 1989, 367-390. 7. Nadkarni, R.A., A review of modern instrumental methods of elemental analysis of petroleum related material, Modern Instrumental Methods of Elemental Analysis of Petroleum Products and Lubricants, ASTM STP 1109, American Society of Testing and Materials, Philadelphia, PA, 1991. 8. Hunt, T.M., Handbook of Wear Debris Analysis and Particle Detection in Liquids, Elsevier Applied Science, London, 1993, 25, and pp. 263. 9. Hunt, J. and Hjelmervik, S., Light-Scattering Laser Sensors for Liquid, Product Note, Met One Inc., Grants Pass, OR, 1991. 10. ASTM, Petroleum Products, Lubricants and Fossil Fuels, Annual Book of ASTM Standards, Vols. 05.01, 05.02, and 05.03, American Society of Testing and Materials, Philadelphia, PA, 1992. 11. Marshall, E.R., Used Oil Analysis—A Vital Part of Maintenance, Lubrication, Vol. 79, No. 2, Texaco, Inc., Beacon, NY, 1993. 12. Skoog, D.A., West, D.M., and Holler, F.J., Fundamentals of Analytical Chemistry, Saunders College Publishing, Ft. Worth, TX, 1988, 418. 13. Lubrizol Petroleum Chemical Company, Diesel Engine Oil Change Intervals, Cleveland, OH, May 1989. 14. Nicolet Analytical Instruments, Theory of FT-IR, Madison WI, 1986, 6.

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15. Toms, A.M., A Preliminary Report on the Evaluation of FTIR for Lubricant Condition and Contamination Determination in Support of Machinery Condition Monitoring in Synthetic Lubricants, Department of Defense, Joint Oil Analysis Program, Pensacola, FL, 1994. 16. Lockwood, RE. and Dally, R., Lubricant Analysis, ASM Handbook, ASM International, Metals Park, OH, Vol. 18,1992,301. 17. Nicolet Instrument Corporation, Used Lubricating Oil Analysis, Madison, WI, 1995, 27.

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Turbine Engine Lubricant 78 Gas Monitoring and Analysis Costandy S. Saba CONTENTS Introduction.........................................................................................................................................................915 Physical and Chemical Properties....................................................................................................................915 Electrical Methods..............................................................................................................................................917 Electrochemical Methods.................................................................................................................................919 Thermal Methods...............................................................................................................................................920 Chemical Methods..............................................................................................................................................920 Future Devices and Techniques for Lubricant Monitoring......................................................................920 References.............................................................................................................................................................921

INTRODUCTION Lubricant condition monitoring is a technique for determining periodically the condition of the lubricant or changes occurring within the lubricant chemistry. For properly running engines, the physical properties of the oil remain fairly constant, since the rate of lubricant degradation is small and some volatile degradation products are lost. When abnormal operating conditions occur which increase aeration rate of the oil and/or oil temperature, the rate of oil degradation increases and the physical and chemical properties of the oil change. Of the numerous physical properties changed by oil degradation, the electrochemical properties are commonly most suited for routine analysis. As an example, the complete oil breakdown rate analyzer measures an electrochemical property ofthe oils which correlates well with conductivity.1-3 By detecting the rapid changes in the electrochemical properties of used oil samples, the Air Force was successful in identifying many abnormally operating engines in the early 1980s.1 While a relationship exists between the degree of oxidative degradation and the electrochemical properties of a lubricant,4 electrochemical measurement of used oils is complicated by the effects of different basestocks and additives. For reliable measurements, one has to stay within the lubricant type and formulation. Along with the comparison in Table 1 of monitoring devices and techniques for used oil, descriptions follow of diagnostic methods based on physical and chemical properties, electrical and electrochemical methods, spectrophotometry, and thermal and chemical stressing.

PHYSICAL AND CHEMICAL PROPERTIES Recent analytical practices for determining the condition of used lubricants have focused primarily on the determination of total acid number (TAN), viscosity, and contaminants (fuel, water, solids). In lubrication systems, lubricants can degrade to produce lower or higher molecular weight compounds Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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which either volatilize and escape out of the engine or remain in the bulk oil, respectively. As a result, the viscosity of the oil will change. According to the MIL-L-7808, 3-cSt lubricant specification, the change in viscosity for laboratory-stressed lubricants should not be more than -5 to +15% after 96 hours of corrosion and oxidation testing at 175°C (Federal Test Method 7915307.1). In real service, a ±15% limit is common.5 Total acid and base numbers are the most commonly monitored chemical properties in used oils. Total base number (TBN) is more of a concern in internal combustion engines where generally over-based mineral oils are used, while total acid number would be for gas turbine engines where ester base lubricants are used. MIL-L-7808 specification requires a change of no more than 2 mg KOH/g for a laboratory-stressed lubricant using corrosion and oxidation testing at 175°C for 96 h, while in service a TAN of 1.0 is a warning limit.5 TAN limits have been used as a criterion for oil changes while rapid rate changes have indicated abnormal conditions. Colorimetric “kits” have been developed for monitoring lubricants based upon changes in TAN which reduce test time but still require disposal of test solutions and impair the accuracy of TAN measurements. Under thermal stressing, deposit-forming tendency of lubricating oils is an important characteristic used to assess lubricant performance.6 Even though total solid determination is a part of used automotive oil testing, gas turbine lubricants are not routinely tested for solids except for cases where the lubricant is visibly black (resulting from lubricant coking) or contaminated by Copyright © 1997 CRC Press, LLC.

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ingestion of debris such as volcanic ash. Similarly, there is no need to analyze for fuel and water contamination in oils from Air Force engines. Fuel contamination can occur from leaks developed in fuel lines, but this occurrence is rare. Water concentration in fresh ester-based lubricant is on the order of 500 ppm. This level is maintained also in used lubricants. While Air Force engines have experienced essentially no corrosion problems due to water build-up, the combination of water- and ester-based lubricant can cause corrosion in systems where magnesium alloy is used. Recently, electric generators on C-17s have failed due to magnesium corrosion. Water level monitoring in used lubricants is more appropriate for equipment aboard ships, steam turbines, or machinery operating in humid environment. Table 2 provides limiting values of lubricant properties measured by various test methods. These values are indicative of lubricant performance under certain test conditions. In order for a lubricant to be qualified for field use, its performance must be within these limits of specification. Using viscosity, acid number, and evaporation loss limiting values (hours) in a laboratory oxidation test, one can assign a ranking for qualifying certain oils. For example, Figure 1 ranks 15 lubricants based on time to reach limiting values, measured in hours, set by the specification of lubricant performance (Table 2). One can easily determine the longest life lubricants, those clustered in the far upper right hand corner of the diagram, when three lubricant properties are considered simultaneously.

ELECTRICAL METHODS CONDUCTIVITY Conductivity measurements using a small portable tester were made on a used 4-cSt ester-based lubricant from a gas turbine engine stand. The results listed in Table 3 correlated well with analysis of oil breakdown rate by COBRA.1,2 The changes in viscosities and TAN values were not significant

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FIGURE 1 Ranking of lubricant lifetimes (hours) based on limiting values of 1.5 TAN, 25% viscosity change at 100°C, and 25% weight loss at 210°C, a modified FTM 5307.1 temperature. even after 144 test hours, indicating no appreciable oxidative degradation. However, the basestock degradation seen from the percent basestock remaining study by gas chromatography, a result of the thermal stressing of the lubricant, contributes to the significant increases in COBRA and conductivity values. While increases in conductivity usually indicate oxidative and thermal deterioration of the oils, only thermal stressing was responsible in this example. DIELECTRIC CONSTANT A dielectric constant tester was used to analyze 3- and 4-cSt ester-basestock aviation fluids that had been stressed in laboratory oxidation tests.7 The data were evaluated with respect to TAN and viscosity of the lubricants and displayed meaningful correlation when evaporative loss of the lubricant was minimal. Changes in the dielectric properties of the lubricant and presence of wear metal debris will effectively alter the capacitance of a test cell, thus altering the frequency of the circuit. A new instrument was evaluated to measure the frequency as a function of time.8 This allows the particles to settle down onto the sensor, altering the frequency and producing a negative sloping line. Furthermore, the instrument also measures dielectric decay in the presence of a time-varying magnetic field. The magnetic “ON” state causes a greater rate of migration of magnetic particles relative to the magnetic “OFF” state, which is shown as a digression in the decay rates of their respective frequency curves. Results with this portable sensor correlated well with oil degradation levels and showed excellent sensitivity for magnetic particles less than 5 (µm at concentrations as low as 1 ppm, but poor sensitivity to nonmagnetic metal particles. MICROSENSOR METHODS Of the many very small chemical and physical sensors being developed, solid state microsensors have the potential of being developed into a lubricant condition-monitoring instrument.9 Jarvis

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et al.10,11 have developed TBN and fuel dilution meters which were successfully tested on operating Navy vessels. These meters use surface acoustic wave (SAW) chemical sensor consisting of a piezoelectric substrate surface (e.g., quartz) imbedded with a pair of integrated electrode arrays which act as a transmitter and receiver of the radio frequency wave. Any absorbed material on the surface of the sensor will cause a shift in the applied radio frequency which is related to the mass/density of the material absorbed. The selection of a chemically specific coating is critical in the development of the sensor response to certain vapors.

ELECTROCHEMICAL METHODS CYCLIC VOLTAMMETRY This technique measures the remaining percent of specific species from the current generated by their oxidation or reduction. It can indicate the remaining useful life of a gas turbine engine lubricant from the electrochemical reduction of the original and generated antioxidant species in the oil sample.12,13 It is applicable to the analysis of lubricating oils, hydraulic fluids, and greases and was successfully field tested in analyzing gas turbine engine oils.

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SPECTROPHOTOMETRIC METHODS INFRARED Generally, this IR absorption technique has been used to monitor the degradation of hydrocarbon (mineral oil or synthetic)-based lubricants by measuring specific absorbencies that appear in the lubricant due to degradation products. The most useful spectral regions that can be monitored are those due to water, soot, oxidation and nitration products, fuel contamination, and lubricant additive depletion. The technique has been shown to be applicable to ester-based lubricants where contaminants and degradation products can be identified.14,15 FLUORESCENCE Fluorescence spectroscopy has been investigated for condition monitoring of oxidized polyphenyl ether lubricants.16 Oxidative degradation of the lubricant was shown to induce specific changes in its fluorescence properties. Excitation/emission wavelengths with monitoring potential were identified and found to yield a useful correlation between emission intensity and viscosity increases of the degraded lubricant for two different formulations. Neat thin films of the degraded fluids yield similar correlations.

THERMAL METHODS Differential scanning calorimetry (DSC), differential thermal analysis (DTA), thermogravimetric analysis (TGA), and thermomechanical analysis (TMA) have found many uses in the lubricant field.17 Among these analytical techniques, DSC is the most useful for lubricant monitoring because it can determine oxidative stability.18 DSC measures heat flow of the sample relative to a reference as a function either of temperature or of time at constant temperature. The onset of rapid oxidation is generally accompanied by a rapid heat flow increase, and lubricant stability is found to be proportional to the induction time or temperature. Both DSC and TGA were used to design a rating system for ester base-lubricants based on their volatility, oxidation, and deposition characteristics.19 DSC is often performed at high pressures (200 to 1000 psi) in order to lower the rate of sample evaporation and sharpen the “onset of oxidation” peak.18 High pressure-differential scanning calorimetry and sealed pan-DSC techniques were used to relate onset of reaction times to the remaining useful lives of ester-based lubricants. A limitation of the DSC technique for condition monitoring is its high instrumental cost relative to electrical methods.

CHEMICAL METHODS Chemical stressing methods involve the addition of reactive species that deplete the antioxidants in the lubricant, thus giving a relative measure of the lubricant’s oxidation resistance. A method was developed to measure the free radical trapping capability of hydrocarbon-base lubricants.20 Later, a technique was developed for determining the peroxide decomposing capability of ester-base fluids of gas turbine engines.21

FUTURE DEVICES AND TECHNIQUES FOR LUBRICANT MONITORING Future monitoring of normal mineral and ester-base lubricants will continue to be based upon measured chemical changes such as total acid numbers and physical property changes such as viscosity and electrochemical properties. Monitoring will likely be conducted more and more in conjunction with wear debris analysis. Such monitoring will be conducted primarily for identifying

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abnormal operating engines and not oil change intervals, since most military aircraft engines require no oil changes between scheduled inspections or overhaul. Monitoring techniques for advanced high temperature lubricants will likely be more specialized as dictated by the chemistry and degradation mode. For example, since polyphenyl ether lubricants do not produce acidic breakdown products, monitoring techniques based on total acid number or electrochemical properties would not be suitable, and specific spectroscopy or chromatography techniques will likely be employed.

REFERENCES 1. Smith, H.A., Complete oil–breakdown–rate analyzer (COBRA) for identifying abnormal operating turbine engines, Joint Oil Analysis Program, Int. Symp. Proc., Pensacola, FL, May 17 to 19, 1983, 307. 2. Saba, C.S. et al., Lubricant Evaluation and Performance, Tech. Rep. No. AFWAL–TR–87–2025, Aero Propulsion Laboratory, Wright–Patterson Air Force Base, OH, June 1987. 3. Saba, C.S. et al., Lubricant Evaluation and Performance II, Tech. Rep. No. WL–TR–91–2111, Aero Propulsion and Power Directorate, Wright–Patterson Air Force Base, OH, January 1992. 4. Saba, C.S. et al., Lubricant Evaluation and Performance, Tech. Rep. No. AFWAL–TR–89–2008, Aero Propulsion Laboratory, Wright–Patterson Air Force Base, OH, April 1989. 5. Centers, P.W., Routine analysis of turbine engine lubricants, Joint Oil Analysis Program, Int. Condition Monitoring Conf., Pensacola, FL, November 14 to 18, 1994. 6. Smith, H.A. and Saba, C.S., Lubricant deposition studies using a small test volume, low cost static coker, Lubr. Engr., 44, 983–992, 1988. 7. Keller, M. and Saba, C.S., Monitoring of ester base lubricants by dielectric constant, Lubr Eng., 45(6), 347–351, 1989. 8. Keller, M.A. and Saba, C.S., Lubricant and wear debris monitoring using a time resolved dielectric device, Joint Oil Analysis Program, Int. Condition Monitoring Conf., Pensacola, FL, November 14 to 18, 1994. 9. Ballentine Jr., D.S. and Wohltjen, H., Surface acoustic wave devices for chemical analysis, Anal. Chem., 60, 704A–715A, 1989. 10. Jarvis, N.L., Wohltjen, H., Klusty, M., Gorin, N., Fleck, C, Shay, G., and Smith, A., Solid–state microsensors for lubricant condition monitoring, I. Fuel dilution meter, Lubr. Eng., 50(9), 689–693, 1994. 11. Woltjen, H., Jarvis, L.N., Klusty, M., Gorin, N., Fleck, C, Shay, G., and Smith, A., Solid–state microsensors for lubricant condition monitoring, II. Total base number, Lubr. Eng., 50(11), 861–866, 1994. 12. Kauffman, R.E., Remaining useful life measurements of gas turbine engine oils, diesel engine oils, hydraulic fluids, gear box oils, and greases using cyclic voltammetric methods, Proc. Joint Oil Analysis Program Int. Condition Monitoring Conf., Pensacola, FL, 1992. 13. Kauffman, R.E., Development of a remaining useful life of a lubricant evaluating technique, III. Cyclic voltammetric methods, Lubr. Eng., 45, 709–716, 1989. 14. Toms, A.M., Joint Oil Analysis Program evaluation of Fourier transform infrared (FT–IR) spectroscopy, I. Proc. Joint Oil Analysis Program Int. Condition Monitoring Conf., November 14 to 18, 1994. 15. Toms, A.M., Joint Oil Analysis Program evaluation of Fourier transform infrared (FT-IR) spectroscopy, II. Proc. Joint Oil Analysis Program Int. Condition Monitoring Conf., November 14 to 18, 1994. 16. Keller, M.A. and Saba, C.S., Monitoring of polyphenyl ether lubricants using fluorescence spectroscopy, Appl. Spect., 44(2), 266, 1990. 17. Noel, F. and Cranton, G., Application of thermal analysis to petroleum research, Amer. Lab., 27–50, June 1979. 18. Kauffman, R.E. and Rhine, W.E., Development of remaining useful life of a lubricant evaluation technique, I. Differential scanning calorimetric techniques, Lubr. Eng., 44(2), 154–161, 1988.

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19. Fehrenbacher, L., Schafrik, R., and Macia, J., Improved Thermo–Oxidative–Deposition Screening Tests for Turbine Engine Lubricants, WRDC–TR–89–2133, Aero Propulsion and Power Directorate, Wright-Patterson Air Force Base, OH, November 1989. 20. Mahoney, L. et al., Determination of the antioxidant capacity of new and used lubricants: methods and applications, Ind. Eng. Chem. Prod. Res., 17, 250, 1978. 21. Kauffman, R.E. and Rhine, W.E., Development of a RULLET, II, Colorimetric Method, Lubr. Eng., 44(2), 162–167, 1988.

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79 Automotive Engine Oil Life Factors Shirley E. Schwartz CONTENTS Introduction....................................................................................................................................923 Chemical and Physical Changes to the Engine Oil During Service........................................924 Standard Tests that Simulate Selected Service Conditions.......................................................925 Significance of Oil Analysis Data................................................................................................926 Conclusion......................................................................................................................................926 References.......................................................................................................................................926

INTRODUCTION Engine oil degrades as a consequence of use in an operating vehicle. The mechanism and rate of engine oil degradation are influenced by many factors, including: • Engine • Design • Presence or absence of hot spots1,2 • Range of operating temperatures1,3 • Nature of the ventilation and recirculation system • Volume of engine oil1,2 • Age • New (green)4 • Well broken in4 • Old and worn5 • Problems • Coolant leak5 • Bearing corrosion or failure, causing hot spots and loss of metal5 • Fuel • Gasoline4 • Diesel oil5 • Alternative fuel4 • Engine oil • Base stock • Synthetic3 • Mineral3 • Additive package Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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• Service • Freeway3,4 • High load5 • City/taxi/police4 • Short trip, cold start6

CHEMICAL AND PHYSICAL CHANGES TO THE ENGINE OIL DURING SERVICE Different types of service cause different changes to the engine oil, as follows: • Freeway/easy driving, in which oil retains its properties for a long time • Viscosity • Slight decrease during first few thousand kilometers after an oil change3 due to shear of viscosity index improver • Slight increase after first few thousand kilometers of driving due to evaporation of light ends of oil or onset of oil oxidation3 • Oil acidity and alkalinity largely stable for many thousand kilometers3 • Antioxidant protection stable3 • High-load, high-temperature driving, in which the engine oil stays hot for extended periods of time • Viscosity increase5 • Evaporation of lighter ends of engine oil • Oxidation of engine oil • Magnitude of the effect influenced by the composition of the oil • Oil acidity and alkalinity5 • Acidity increase, alkalinity decrease • Magnitude of the effect influenced by • Composition of the engine oil • Temperature to which the oil rises • Accumulation of insoluble contaminants in engine oil5 • Oil oxidation products • Metals • Wear5 • Increase in oil additive concentration due to evaporation of lighter oil ends • City driving, with extended idle, rich operation, engine oil completely warm • Increase in oil acidity, viscosity, and insoluble contaminants compared to freeway service4 • Decrease in oil alkalinity and oxidative stability compared to freeway service4 • Short-trip, cold-start driving • Significant amount of contamination in engine oil • Fuel (up to 29% observed under extreme conditions)7 • Water (up to 9% observed with gasoline-fueled vehicles) and emulsion formation3 • Acid from partially combusted fuel7 • Effect partially reversed when oil warms6 • Additive drop out to bottom of pan and additive depletion in bulk of oil • Water and emulsion formation a primary cause6 • Effect reversible if water driven off by higher oil temperatures6 • Viscosity • Up to 70% viscosity reduction due to fuel in oil7 • Oil thickening due to gel formation with polar oil contaminants

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• Increased rate of entry of iron into engine oil • Oxidation (rust formation) due to the presence of water7 • Corrosion from acidic oil7 Some of the changes that occur in the engine oil as a consequence of various engine conditions are: • Green (new) engine causes high concentration of metals and silicon in engine oil7 • Bearing problems may cause high concentration of lead, copper, or aluminum in engine oil5,8 • Coolant leaks cause an elevated concentration of coolant components in engine oil9 • Silicon, sodium, glycol often identifiable • Glycol sometimes not detected when present • Positive test for glycol seen with some unused oils • High-load, stop-and-go driving with a Diesel engine creates soot, which enters the engine oil5 By comparing oil analysis results for a given type engine, but different type of service, it is even possible to obtain a quantitative assessment of the severity of one type of service relative to another, as seen in the following examples: • Reaching a 95% reduction in remaining antioxidant protection required • 12,000 km in freeway service • 9,000 km when pulling a 1000-kg trailer in the same service2 • Reaching a 30% increase in viscosity (gasoline-fueled vehicle) required • 16,000 km in freeway service • 8,000 km in city service4 • Reaching 40 parts per million iron in engine oil quired • 7,000 km in freeway service • 1,000 km in extreme short-trip, cold-start driving3 These examples indicate that there are significant differences in the rate of engine oil degradation as a function of type of service.

STANDARD TESTS THAT SIMULATE SELECTED SERVICE CONDITIONS American Society for Testing and Materials (ASTM) standard test methods have been developed to qualify an engine oil for use in an operating vehicle. Each test stresses the engine oil in a specific way. For example, the following tests represent severe examples of three of the types of service listed in the introduction: • Sequence IIIE simulates high-temperature, high-load conditions and assesses an oil’s ability to resist oxidation and thickening.10 • Sequence VE simulates city type driving and determines an oil’s ability to resist the changes that occur when an oil is exposed to the chemicals that form during city service.” • Sequence IID simulates short-trip, cold-start driving and provides a measure of an oil’s ability to inhibit rust and corrosion.12 In such tests, engine measurements and oil analyses indicate whether an engine oil provides acceptable protection to an engine. Information on additional test procedures for such thins as bearing corrosion, oil thickening, foaming tendency, or corrosion inhibition is also available through such organizations as the Copyright © 1997 CRC Press, LLC.

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International Lubricant Standardization and Approval Committee (ILSAC), the Society of Auto-motive Engineers (SAE), Japan Automobile Standards Organization (JASO), and the Coordinating European Council (CEC).

SIGNIFICANCE OF OIL ANALYSIS DATA One can often determine what kind of service an engine has experienced by analyzing the engine oil.9 For example, if the oil viscosity is low and fuel and water are present (the oil may look milky), the vehicle was probably driven in short-trip service. If the oil is thick and tarry, the driver almost certainly drove too long before an oil change, or perhaps there was a coolant leak. If the concentration of engine metals and silicon in the oil seems too high, the oil sample may have been taken during or at the end of the first oil drain. If the oil seems thick and the concentration of oil additive elements seems too high, the vehicle was probably driven under high-load, high-temperature conditions, and the lighter ends of the oil have evaporated off. If lead and copper concentrations seem high, but the concentration of other metals seems normal, there may be a bearing problem.

CONCLUSION The properties of engine oil are modified in specific ways for any given type of service. Analysis of the engine oil provides considerable insight into the severity of service, the presence of any problems, and the durability of the engine. The nature of the change in engine oil properties also may provide clues regarding engine malfunction.

REFERENCES 1. Schwartz, S. E., A model for the loss of oxidative stability of engine oil during long-trip service, I. Vehicle measurements, STLE Tribol. Trans., 35(2), 307–315, 1992. 2. Schwartz, S. E., A model for the loss of oxidative stability of engine oil during long-trip service, II. Theoretical considerations, STLE Tribol. Trans., 35(2), 235–244, 1992. 3. Younggren, P. J. and Schwartz, S. E., The effects of trip length and oil type (synthetic versus mineral oil) on engine-oil degradation in a driving test of a vehicle with a 5.7L V-8 engine, SAE Pap., No. 932838, Society of Automotive Engineers, Warrendale, PA, 1993. 4. Schwartz, S. E. and Mettrick, C. J., Mechanisms of engine wear and engine oil degradation in vehicles using M85 or gasoline, SAE Pap., No. 942027, 1994. 5. Schilling, A., Automobile Engine Lubrication, Scientific Publications (G. B.), Broseley, Shropshire, England, 1972. 6. Schwartz, S. E., Observations through a transparent oil pan during cold-start, short-trip service, SAE Pap., No. 912387, 1991. 7. Schwartz, S. E., A comparison of engine oil viscosity, emulsion formation, and chemical changes for M85 and gasoline-fueled vehicles in short-trip service, SAE Pap., No. 922297, 1992. 8. Zuidema, H. H., Bearing lubrication, in The Performance of Lubricating Oils, Reinhold Publishing, New York, 1959, 89–113. 9. Smolenski, D. J. and Schwartz, S. E., Automotive engine-oil condition monitoring, in CRC Handbook of L ubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 17–32. 10. Standard Test method for Evaluation of Automotive Engine Oils in the Sequence IIIE SparkIgnition Engine, ASTM D 5533, Vol. 05.03, American Society for Testing and Materials, Philadelphia, 1996, 659–765. 11. Standard Test Method for Evaluation of Automotive Engine Oils for Inhibition of Deposit Formation and Wear in a Spark-Ignition Internal Combustion Engine Fueled with Gasoline and Operated Under Low-Temperature, Light-Duty Conditions, ASTM D 5302, Vol. 05.03, 1996, 429–540. 12. Test Method for Evaluation of Automotive Engine Oils for Inhibition of Rusting (Sequence IID), ASTM D 5844, Vol. 05.03, ASTM, West Conshohocken, PA, 1995.

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Automotive Engine Oil Condition Monitoring Donald J. Smolenski and Shirley E. Schwartz

Analysis of used engine oils can often provide valuable information about: • • • • •

Extent of engine oil degradation Type of service the engine has experienced Condition of the engine Nature and extent of possible engine problems Probable cause of engine failure.

In the tables, the following information is presented: • Test method • ASTM (American Society for Testing and Materials) designation (if an ASTM method is available) • Warning limits (with references) • Significance of test results • Effect of driving conditions on the interpretation of test results • Related analyses which may support or refute the conclusions drawn from a given test result. The test methods listed are not a complete compilation of all pertinent methods available, but include methods which have been found useful to investigators. Often related methods are listed in the ASTM procedure.

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REFERENCES 1. Younggren, P. J. and Schwartz, S. E., The effects of trip length and oil type (synthetic versus mineral oil) on engine damage and engine-oil degradation in a driving test of a vehicle with a 5.7L engine, SAE (Society of Automotive Engineers) Pap., No. 932838, 1993. 2. Schilling, A., Automobile Engine Lubrication, Scientific Publications (G.B.), Broseley, Shropshire, England, 1972. 3. Rodgers, J. J. and Kabel, R. H., A Revised Sequence IIIC Engine Oil Test, General Motors Research Laboratories Publ. 2611, Warren, MI, 1978. 4. Engine Oil Licensing and Certification System, Appendix K–2, API Publ. No. 1509, 12th ed., American Petroleum Institute, Washington, D.C., January 1993, 51. 5. Testing Used Engine Oils, Chevron Research Bulletin, Richmond, CA, 1983. 6. Engine Oil Viscosity Classification, SAE J300, Society of Automotive Engineers, Warrendale, PA, March 31, 1993. 7. Firey, J. C., Newcomb, J. C., Niemann, J. F., and Sugges, P. R., Studies of the effects of water on gasoline engine wear at low temperature, Wear, 10, 33, 1967. 8. Dyson, A., Richards, L. J., and Williams, K. R., Diesel engine lubricants: their selection and utilization with particular reference to oil alkalinity, Proc. Inst. Mech. Eng., 171, 717, 1957. 9. Schwartz, S. E. and Smolenski, D. J., Development of an automatic engine oil-change indicator system, SAE Pap., No. 870403, 1987. 10. Schwartz, S. E., A comparison of engine oil viscosity, emulsion formation, and chemical changes for M85 and gasoline-fueled vehicles in short-trip service, SAE Pap., No. 922297, 1992. 11. Smolenski, D. J. and Kabel, R. H., Effect of engine oil zinc dithiophosphate (ZDP) additive type on cam and lifter wear in taxi service, SAE Pap., No. 831760, 1983. 12. Schwartz, S. E. and Mettrick, C. J., Mechanisms of engine wear and engine oil degradation in vehicles using M85 or gasoline, SAE Pap., No. 942027, 1994. 13. Mettrick, C. J. and Schwartz, S. E., A study of the effects of extended oil-drain periods on engine oil degradation and engine durability in a fuel-flexible vehicle, in Proc.: Vol. 1, Tenth International Symposium on Alcohol Fuels, November 7 to 10, 1993, 83–91. Also available in General Motors Research and Development Publ. R&D-8035, Warren, MI, 1994.

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81 Diesel Engine Used Oil Analysis Jack Poley CONTENTS Introduction....................................................................................................................................935 Sampling Guidelines......................................................................................................................935 Tests for Used Diesel Lubricants.................................................................................................936 Wear Particles..................................................................................................................................936 Tables...............................................................................................................................................936

INTRODUCTION Used lubricant analysis as a maintenance aid is now a mature field, particularly for diesel engine oils. In about 1948, railroad companies began to utilize their own laboratories to analyze oils for wear debris in addition to commonly requested physical and chemical tests. Since that time, lubricant analysis has expanded to trucking, agriculture, construction, industry, marine applications, oil and gas operations, and aviation. Today’s testing is faster, less expensive, and better than was previously the case. In addition to improvements in instrumentation, the advent of the computer permeates every aspect of the analytical process: logging in samples, analyzing and storing data, predicting trends, and even designing the final report. In the future, it is probable that routine evaluation of lubricants may be conducted in situ by using on-board sensor technology. On the other hand, it is also anticipated that lubricant sampling and analysis technology will continue to improve, since there will always be a need for more sophisticated testing whenever on-board sensors have detected an anomaly.

SAMPLING GUIDELINES It is relatively easy to obtain engine oil samples from most diesel engines. However, a few fundamentals should be observed to maximize the quality of the samples: • Samples are taken • From an engine that is thoroughly warm. This ensures a well-mixed, representative sample. • From the same location every time. • Samples are not taken Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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• From gauge lines or other static locations. Samples taken in this fashion may represent very old lubricant from drain cycles many times previous. • Immediately after the filter. The filter removes a lot of representative debris or “evidence.”

TESTS FOR USED DIESEL LUBRICANTS A thorough lubricant analysis procedure includes several tests. No single test provides sufficient information. In the enclosed tables, recommended tests for diesel engine applications are listed, along with typical results and limits.

WEAR PARTICLES Wear particles are a special classification of debris. Those that are very small (less than two or three micrometers, µm) are routinely detected via spectrometric analysis, which is the most common type of lubricant analysis. Many techniques for analysis or particle counting depend on light passing through the oil sample. However, if soot from the fuel is present in the oil, the soot may block the light required for analysis. Alternatives to light transmission techniques are becoming available. Particles appreciably larger than a few micrometers are more readily analyzed by alternative methods. For diesel engine oils, ferrography represents an excellent approach for counting particles, since one can determine the size, shape, and metallurgical composition of the particles. This combination of shape and metal identification makes ferrography a powerful decision–making tool whenever engine inspection or routine oil analysis indicate that the engine might have a serious problem. Simple particle counts which do not identify the composition or structure of the particles are not nearly as useful.

TABLES The following tables describe the analytical measurements that are typically done on diesel engine oil samples. Interpretation of analytical results may depend on the particular application, type of service, type of oil used, type of engine, and the nature of any damage that may have occurred during the sampling period. Thus, the interpretations presented in the tables are general guidelines that will be correct for most applications, but may not be appropriate for some specialized situations.

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Rotating Machinery Vibration and Condition Monitoring William D. Marscher

CONTENTS The Role of Vibration Analysis in Condition Monitoring..................................................................944 Avoidance of Fluid-Induced Vibration Problems................................................................................946 Balance Criteria............................................................................................................................................946 Alignment Criteria.......................................................................................................................................947 Machinery and Piping Support.................................................................................................................947 Resonance.....................................................................................................................................................948 Rotordynamic Instability............................................................................................................................948 Vibration Specifications.............................................................................................................................948 Troubleshooting...........................................................................................................................................950 Recognizing Vibration Problems.............................................................................................................950 References.....................................................................................................................................................955

THE ROLE OF VIBRATION ANALYSIS IN CONDITION MONITORING Rotating machinery can experience reliability problems that are evidenced by increased vibration at certain frequencies. These reliability problems are usually associated with either fatigue (development of cracks in components) or tribological degradation (increased friction and/or wear). Figure 1 illustrates typical mechanical durability issues in fluid dynamic rotating machinery. Vibration probes and associated process and lubricant transducers are typically installed at several or all of the locations shown in Figure 2. Vibration test data taken from the vibration probes, to help determine the current and projected mechanical health of rotating machinery, are generally plotted in three different forms: 1. “Signature” or spectrum plots of vibration amplitude vs. frequency. When combined with a frequency plot of the phase lag angle between application of the exciting force and the resulting vibrational motion, this is called a “Bode plot.” 2. Plots of vibration amplitude vs. time, similar to a typical oscilloscope trace. 3. “Orbit” plots of the shaft’s centerline radial motion over one period of vibration, or one shaft revolution.

The amplitude scales are usually linear, but are sometimes given in decibels (dB). dB is a base 10 logarithmic representation used to enhance the resolution of low-level responses in the “noise floor” of a spectrum, such as fluid excitation forces or natural frequency peaks. Whether linear or Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 Typical internal failure types in turbomachinery.

FIGURE 2 Turbomachine well instrumented according to current practice. dB, the amplitude in most spectrum analyzers usually indicates root mean square (RMS) values, which are a factor of 0.707 times peak values.

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AVOIDANCE OF FLUID-INDUCED VIBRATION PROBLEMS Whenever possible, a fluid machine should be operated in the vicinity of its design point. For all types of turbomachinery, vibration, bearing load, and associated eccentricity typically increase rather than (as intuition would suggest) decrease when running at low flows. For example, if turbomachines are operated far enough below their design point, stalling of either the rotor blades or diffuser vanes may occur, accompanied by large excitation forces at frequencies of about 20% and 80% of running speed, usually associated with a phenomenon called “rotating stall.” At still lower flows in fans and compressors, violent flow pulsing in the compressor and its system may occur, which is commonly known as “surge.” Both surge and stall are generally associated with poor angle of attack between incoming flow and the blade or vane angles built into the various turbomachine components. This often causes fluid “dead” zones, with eddies developing in the resulting wake, possibly resulting in “recirculation” heading upstream of the component. This in turn causes severe shaking of the rotor and of stator assemblies, particularly if the frequency at which the eddies propagate coincide with a component natural frequency. For example, in pumps, recirculation can cause severe vibration levels if the below-running-speed excitations coincide in frequency with a rotor critical speed, and sometimes causes cavitation damage on the pressure side of impeller vanes. In compressors, the stalling from this recirculating flow is often a precursor to surge, which is a violent and potentially destructive transient shaking of the entire machine due to temporary mismatch of the flow capacity within a stage’s aerodynamic passages or between stages. For centrifugal turbomachinery, it is possible to trade somewhat diminished efficiency for lower blade or vane pass forces by increasing the clearance between the rotor blades and the volute tongue or diffuser vanes. In centrifugal pumps, this is the so-called “B-Gap.” Excessive vane passing vibration becomes more likely if the ratio of diametral blade/vane gap vs. impeller diameter is less than about 5%. However, excessive gap can encourage part load discharge recirculation. A sound approach is to follow manufacturer guidance on the setting of rotor/stator gaps. For all types of pumps, in order to minimize hydraulic forces in general, it is important to operate the pump with sufficient net positive suction head (NPSH), the effective absolute static minus vapor pressure at the pump suction. If inlet static pressure drops below the pumped fluid vapor pressure, which it is prone to do at the low pressure or “suction” side of impeller vanes near the leading edge, then local boiling or “cavitation” will occur. Microscopic bubbles form and then collapse abruptly as they enter zones of significantly higher pressure. These imploding bubbles shock nearby metal, and often cause serious erosion damage to the impeller. They result in an accompanying low frequency as well as very high frequency shaking of the pump. This shaking may lead to unusually high vibration of various natural frequencies. Maintenance of sufficiently high suction pressure or “NPSHA” (NPSH available) is required to avoid cavitation, at least equal to the “NPSHR” (NPSH required) and preferably at least twice this value. This implies that the designer should strive for low inlet velocities, and therefore large diameter suction pipes. Any valves, elbows, or pipe reducers to match this pipe to the pump suction flange should preferably be well upstream of the pump flange (at least five diameters).

BALANCE CRITERIA Adequately balance all rotating components that have a diameter greater than 1/3 of the disk or impeller diameter. Two-plane balance all components that have a length of greater than 2/3 their diameter for machines operating at 3600 rpm and below and greater than 1/4 their diameter for faster machinery. The entire rotor system should be check-balanced once assembled. The most typical form of balance standard, which is based on induced imbalance loads in a stiff shaft, is

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where:

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z is a constant equal (in SI units) to 0.1 to 0.4 depending upon machinery type (generally higher speed machinery should have a proportionately lower value, with 0.4 being ideal for a speed of 1800 rpm) e is the imbalance, kg-mm W is the mass of the balanced component, kg N is the peak operational speed, Hz.

Note that in English units z is between 4 and 16, e is in oz-in., W is weight in lb, and N is speed in rpm. Different balance grades exist for different types of equipment, and z values significantly less than the range specified above may be used for very high speed or high power density equipment. In lower speed or lower performance machinery, on the other hand, looseness in rotating component fits should be limited such that the mass of the component times the radial clearance does not violate the above criterion.

ALIGNMENT CRITERIA The purpose of alignment is to make the shaft centerlines of the equipment and its driver nearly concentric and parallel. To specify adequate coupled shaft alignment for a rigid or single engagement flexible coupling, a good rule-of-thumb is to maintain 50 µm (2 mils) or less concentricity between the opposing stub shaft centerlines. This criterion may be loosened by an additional 1 mm/m (about 1 mil/in.) of length between coupling hubs for flexible couplings of the double engagement or “spacer” type. Maximum permissible angle between shafts at the point of coupling engagement is about 5 minutes (this is automatically maintained if the offset spec is maintained). Coupling manufacturers often quote larger amounts of offset and parallel misalignment than the above values, but these quotes do not account for the additional equipment vibration that misalignment might cause, and hence focuses only on coupling survivability, not equipment survivability. Alignment is best determined by the reverse dial indicator method or its modern optical equivalent, as detailed by Karassik et al.1 Steps must be taken to ensure that the alignment criteria also apply to the “hot” alignment, i.e., with the rotating machine running after warm-up. Also, be sure to compensate for the run-out of any indicating surface vs. the stub shaft centerline, and for gauge support cantilever sag.

MACHINERY AND PIPING SUPPORT The mass of machinery foundations should be about 10 times that of the machine itself. Machinery baseplates should overlap foundation pads by at least half of the floor plus pad thickness. The foundation should have a low center of gravity vs. the spread of its foundation bolts, to provide sufficient base moment support or “footing.” The floor stiffness can influence the machine’s structural natural frequencies, which the manufacturer calculates based upon an infinitely stiff floor. The component masses and stiffnesses are available from the manufacturers upon request, so that structural natural frequencies can be calculated. These calculations are best done with the finite element method. Both the inlet and discharge piping flanges and the opposing fluid machinery nozzles should be properly supported. The use of unrestrained pressure-bearing “expansion” or “flexible” joints at machine nozzles can be tolerated for piping less than 6 inches in diameter, or for piping in which the contained pressure times the nozzle cross-sectional area is within the manufacturer’s nozzle load limits. Such joints relieve any piping thermal expansion from loading up the equipment nozzles. However, because they do not allow the piping to absorb the cross-sectional nozzle fluid pressure thrust (i.e., the pressure times the open area), larger nozzles and nozzles which do not meet machinery manufacturer stated limits must be restrained by axial tie-rods or flange bulkheads firmly anchored in the foundation. Otherwise, pressure forces perpendicular to the nozzle opening can

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put severe loads on the machine’s casing, leading to severe misalignment, possible rubs, and even possible cracking of the nozzle at its neck.

RESONANCE Even if mechanical and fluid dynamic forces are properly controlled, it is still possible that natural frequency resonance may cause vibration to be excessive. To avoid this problem, no natural frequencies of the rotor or structure should be close to excitation frequencies, such as lx, 2x, rotor blade number, or stator vane number times running speed. This is accomplished in the design stage using appropriate computer models, such as finite element analysis for static structures, and specialized rotordynamic computer programs for rotor systems. In such models, the rotating machine component drawings must be obtained from the machinery manufacturers (including dimensions, clearances, and weights of components), and must be available from the system designer for the machine base (including ribs), floor (including subfloor details and supports), and piping. Simple manual calculations are not likely to be sufficiently accurate to avoid resonances, because of the complexity of most rotating machinery and their installations. Also, individual component natural frequencies are not necessarily relevant to the final assembled system natural frequencies, because of the leverage and altered effective stiffness and effective mass of the interlinked components.

ROTORDYNAMIC INSTABILITY The possibility of “rotordynamic instability” due to lateral motion needs to be checked for in centrifugal and axial machinery, particularly in large compressors, fans, and certain motors in which crosscoupling (in which the forces induced by rotor motion are perpendicular to the motion) is often large enough to possibly overwhelm the damping. Rotordynamic instability refers to phenomena whereby the rotor and its system of reactive support forces, particularly the cross-coupling force acting perpendicular to displacement, are able to get out of phase with each other and become self-excited, leading to potentially catastrophic vibration levels, even if the original excitation forces are quite low. Rotors are particularly prone to this phenomenon if the logarithmic decrement (“log dec,” approximately equal to the critical damping ratio times 2 pi) is less than 0.2 for rotor bending natural frequencies less than half running speed. The characteristic of rotordynamic instability, sometimes called “shaft whip,” is a whirling at about half running speed, beginning when running speed exceeds twice the first bending natural frequency of the shaft. This causes a double loop orbit, such as shown in Figure 3. If an unstable machine is encountered, typical design modifications which reduce the tendency to rotor dynamic instability involve bearing changes. The type of bearing most likely to participate in instability problems is the plain journal bearing, which has very high cross-coupling, although it also has high beneficial damping. Bearings which discourage whirling lubricant flow, such as lemon bore bearings, tend to decrease cross-coupling. The most effective bearings in this regard are axially grooved and tilting pad bearings.

VIBRATION SPECIFICATIONS Injury to machines due to excessive rotor vibration takes the form of wear or fatigue damage to the internal components, such as bearings, annular seals, mechanical seals, and the shaft. The majority of these damage mechanisms depend upon the maximum displacement caused by the vibration. However, because rotor stress is limited by material strength, and the stress-to-strength ratio increases with the square of the rotor outer diameter “tip speed,” then as machines are made faster, they must on the average become proportionately smaller. Therefore, as machines increase in speed, on average the amount of vibration displacement they can tolerate decreases proportional to the speed increase.

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FIGURE 3 “Half-speed” vibration orbits. For this reason, the allowable running speed vibration velocity (which equals a constant times displacement times running speed) is roughly constant regardless of the running speed of the machine. Machine survival vs. failure data published by insurance companies and machinery “users” support the use of constant vibration velocity vs. speed as an acceptance criterion in assessing vibration severity. This was first presented 60 years ago by Rathbone,2 and 30 years later was confirmed by Blake,3 Baxter and Bernhardt4 and Hancock.5 Since that time, little additional research has been performed or published concerning this issue. However, the raw information in the quoted references was based on measurement equipment that could not distinguish between various frequency components. Therefore, vibration severity could be plotted only as unfiltered (i.e., total vibration, including all frequencies) displacement readings vs. machine running speed, not filtered (i.e., individual values at specific frequencies) velocity values vs. frequency. Unfortunately, many specifications (with the notable exceptions of API 610,6 and the ANSI/Hydraulic Institute Standards 15th edition7) have assumed that the original data could be interpreted as velocity vs. frequency (since frequency has the same units as speed), and have presented their specifications on this basis. Be very cautious in using such specifications, since they will tend to overlook instability and hydraulic problems causing rubbing at low frequencies, and may require unnecessarily small vibration displacements at high frequencies, causing rejection of good equipment at the expense of all involved, as discussed by Marscher.8 Typical specifications used to establish vibration test types, measurement locations, and acceptance criteria, are the ANSI/Hydraulic Institute Standards,7 API 610,6 and various ISO machinery vibration specifications such as ISO 2372.9 These specifications require the monitoring of bearing housing vibration, using an accelerometer or velocity probe in three perpendicular directions at the Copyright © 1997 CRC Press, LLC.

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location of each of the bearings in the drive train. The acceleration readings are typically integrated to obtain vibration displacement and/or velocity values, which are the terms in which the vibration acceptability criteria are typically given. Occasionally, the specifications give procedures for installation and evaluation of shaft proximity probes, mounted in some bearing housings to measure shaft vs. housing displacement, particularly in higher speed or high power density machinery. Acceptability criteria vary. A generally conservative approach consistent with all of the quoted specifications is to require that vibration does not exceed any of the following three criteria, at any frequency: 2.0 mils displacement peak-to-peak, 0.25 in./s zero-to-peak (or just “peak”) velocity, and 1.0 g peak acceleration. However, keep in mind that the best specification levels are dependent upon the specific machine, its service, and the closeness of the operating point to the design point. Other issues important to consider in setting quantitative specification acceptance levels are listed in Table 1.

TROUBLESHOOTING Sometimes, troubleshooting using detailed vibration testing is required. The most common type, “signature analysis,” generally uses accelerometer output sent to a fast Fourier transform (FFT) analyzer to document the amount of vibration at each frequency within a tested range. Typically this range is from several Hz to beyond the machine’s blade or vane pass frequency. The frequencies at which most of the vibration is occurring, and the locations where the vibration is the greatest, are used as clues to determine the cause of the vibration. A list of the most typical vibration excitation frequencies and associated possible causes is given in Table 2. Often vibration testing ends here. However, it is recommended that vibration testing include experimental modal analysis (EMA). EMA involves artificially exciting a machine or structure such as with an impact hammer, preferably while the machinery is running, so that all bearing stiffnesses are representative, as discussed by Marscher.10 The purpose of EMA is to determine the natural frequencies of a rotating machine, its rotor system, and attached system components. These can be compared to excitation frequencies to determine whether resonance is occurring, and whether all equipment and the associated system and supporting structures have adequate separation margins between the excitation frequencies and natural frequencies.

RECOGNIZING VIBRATION PROBLEMS The great majority of machinery vibration problems can be solved by re-balancing of the rotor assembly, alignment of the coupling (especially when the system is warmed up and operating) e.g., per Dodd,11 and/or running all fluid machines within the bounds of their specified pressure vs. flow

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curves. Remaining vibration problems are generally due to a resonance of a natural frequency. During resonance, vibrations can exceed internal clearances, or excessive bearing loads can occur, even if forces such as imbalance or misalignment are within normally acceptable limits. In performing vibration troubleshooting, a generalized chart such as Table 3 matching symptoms to possible causes can be useful for many typical or simple problems. This chart is not all inclusive and is in the order of the frequency value observed, not in order of likelihood or importance to reliability. A pictorial illustration of some of the symptoms described above, matched to the original problem and its ramifications, is given in Figure 4.12

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FIGURE 4 Important vibration problems in turbomachinery. (From CRC H andbook of L ubrication and Tribology, Vol. 3, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1994, 65-66. With permission.)

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FIGURE 4 (Continued)

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REFERENCES 1. Karassik, I.J., Krutzsch, W., Fraser, W., and Messina, J., Eds., Pump Handbook, McGraw-Hill, 1984, 288–301. 2. Rathbone, T., Vibration tolerance. Power Plant Eng., Vol. 43, November 1939. 3. Blake, M., New vibration standards for maintenance, Hydrocarbon Processing and Petroleum Refining, January 1964. 4. Baxter, R.L. and Bernhardt, D.L., Vibration Tolerances for Industry. ASME 67–PET–14, American Society of Mechanical Engineering, New York, 1967. 5. Hancock, W.P., How to control pump vibration, Hydrocarbon Processing, March 1974. 6. API 610, 8th ed., American Petroleum Institute, Washington D.C., 1995. 7. ANSI Standards, 15th ed., Hydraulic Institute, Parsippany NJ, 1995. 8. Marscher, W., The relationship between rotor system tribology and appropriate vibration specifications for centrifugal pumps, Proc. Int. Mech. Eng. 3rd Eur. Congr. Fluid Machinery for the Oil and Petrochemical Industries, The Hague, Netherlands, May 1987. 9. ISO 2372, Mechanical Vibration of Machines, International Organization for Standardization, Geneva, 1974. 10. Marscher, W.D., Determination of rotor critical speeds during operation through use of modal analysis, Proc. ASME 1986 WAM Symp. on Troubleshooting Methods and Technology, Anaheim, CA, December 1986. 11. Dodd, V.R., Total Alignment, Petroleum Publishing, Tulsa, OK, 1974. 12. Marscher, W.D., Rotating machinery vibration testing, condition monitoring, and predictive maintenance, in CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E.R., Ed., CRC Press, Boca Raton, FL, 1994, 43–70.

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and Roller Bearing 83 Ball Troubleshooting* Thomas Jendzurski TYPICAL SYMPTOMS FOR IMPENDING BEARING FAILURE Bearings which are operating improperly usually exhibit identifiable symptoms. This section organizes these symptoms and provides useful hints to help prevent reoccurrence. The following are the seven most common symptoms of bearing trouble (Table 1): A. Overheated bearing B. Noisy bearing C. Replacements are too frequent D. Vibration E. Unsatisfactory equipment performance F. Bearing fit loose on the shaft G. Shaft is difficult to turn Under each symptom in the following lists are given a number of possible reasons for the condition, practical solutions which may be possible, and a visual illustration of the nature of the problem. Depending on the degree of bearing damage, many misleading symptoms may be present. In most cases these misleading factors are the result of secondary damage. To effectively troubleshoot for the underlying problem, it is necessary to analyze those symptoms which are first observed in the application. As a further aid in troubleshooting, a photographic “Bearing Failure Guide” (Table 2) follows the tabulation of symptoms. While many bearing failures can be examined satisfactorily with a magnifying glass, a zoom lens binocular microscope is preferable. Starting with about 5× magnification to minimize distortion created by the curvature of bearing surfaces, details in a damaged area can then be examined more closely at higher magnifications.

* Text and figures for this chapter are published with permission of SKF Bearing Services Co., King of Prussia, PA.

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958

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959

960

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961

962

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963

964

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965

966

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967

968

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969

970

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971

972

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973

974

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975

976

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977

978

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979

980

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981

982

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983

984

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84 Gear Distress and Failure Modes Courtesy of the Falk Corporation CONTENTS Surface Fatigue................................................................................................................................986 Wear..................................................................................................................................................993 Plastic Flow...................................................................................................................................1002 Breakage........................................................................................................................................1004 Failures Associated with Processing.........................................................................................1006

Distress or failure of gears may be classified into four categories: 1. 2. 3. 4.

Surface fatigue (pitting), Wear, Plastic flow, Breakage.

The appearance of the various distress and failure modes can differ between gears that have through-hardened teeth and those that have surface-hardened teeth. These differences result from the different physical characteristics and properties and from the residual stress characteristics associated with the surface-hardened gearing. Where appropriate, examples of distress of both through- and surface-hardened gears are shown and discussed.

SURFACE FATIGUE Surface fatigue is the failure of a material as a result of repeated surface or subsurface stresses beyond the endurance limit of the material. Figure 1 indicates the theoretical mutual Hertzian stresses occurring when a gear and pinion mesh. There are compressive stresses at the surface and unidirectional and bidirectional subsurface shear stresses. Figure 2 indicates the magnitude of these stresses. PITTING Pitting is a form of surface fatigue which may occur soon after operation begins and may be of three types: 1. Initial (corrective) 2. Destructive 3. Normal

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FIGURE 1 Hertzian stresses.

FIGURE 2 Variation of stress below surface. Initial pitting is caused by local areas of high stress due to uneven surfaces on the gear tooth. This type pitting can develop within a relatively short time, reach a maximum and with continued service polish to a lesser severity. Initial pitting shown in Figure 3 usually occurs in a narrow band at the pitchline or just slightly below the pitchline. It is most prominent with through-hardened gears. Shape of a classical pit is shown in Figure 4. It appears as an arrowhead pointing in the direction of oncoming contact. Starting at the surface of the tip of the arrowhead, the fracture proceeded inward at a shallow angle to the surface. Simultaneously, the crack broadened, forming the arrowhead. The

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FIGURE 3 Initial (corrective) pitting.

FIGURE 4 Initial (corrective) pitting (enlarged photograph). back side of the pit has a steep side. Although there were several large pits on the tooth surface, this pitting was corrective, since it progressed no further with continued operation. For most through-hardened industrial-type gears, initial pitting is considered normal and no remedial action is required. Where necessary, initial pitting can be reduced by special tooth-finishing means and sometimes by a careful break-in at reduced loads and speeds. In some special critical applications, teeth are copper or silver plated to prevent or reduce initial pitting. Destructive or progressive pitting, on the other hand, usually starts below the pitch line, in the dedendum portion of the tooth and progressively increases in both the size and number of pits until the surface is destroyed. While destructive pitting can appear to be as severe as corrective pitting at the beginning of operation, as time goes on the severity of destructive pitting sharply increases and far surpasses the severity of corrective pitting as shown in Figure 5. Figure 6 illustrates this type of pitting on a through-hardened gear. Figure 7 shows destructive pitting of a surface-hardened gear. Destructive pitting usually results from surface overload conditions that are not alleviated by initial pitting. If tooth surface hardness is within specified values, system overloads are usually the

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FIGURE 5 Pitting severity.

FIGURE 6 Destructive (progressive) pitting, through-hardened gears.

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FIGURE 7 Destructive (progressive) pitting, surface-hardened gears. cause of such pitting. To see a finely pitted gear with several large pits is no cause for alarm since it can be of a corrective nature. Normal dedendum pitting wear of fully loaded through-hardened gears manifests itself as small or modest size pits, covering the entire dedendum portion of the tooth flanks. Continued operation results in pit rims being worn away with virtually no further pitting occurring. Figure 8 shows the tooth appearance in the pitting phase, prior to pit rim wear. Figure 9 illustrates a dedendum pitted gear tooth after pit rims have been worn away.

FIGURE 8 Normal dedendum wear. Copyright © 1997 CRC Press, LLC.

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FIGURE 9 Normal dedendum wear.

FIGURE 10 Tooth surface microcracks. Dedendum pitting results when loads are at or close to maximum allowable surface-loading values. The dedendums are most vulnerable to this phenomenon because of the preferential orientation of the surface microcracks along the tooth profile. Figure 10 illustrates this. The orientation of the cracks in the dedendum of both pinion and gear are such that oil is readily trapped in them as the contact rolls over the surface openings. These then propagate rapidly into pits by hydraulic pressure. In the addendum, the oil is forced out of the microcracks before the contact progresses far enough to seal the surface openings off; hence, hydraulic propagations of the crack are almost nil and few pits are formed in this region. At loadings currently used for industrial surface hardened gears, pitting is much less prevalent than with through-hardened gears. When it does occur, the appearance may be similar to that of through-hardened gears. Micro-pitting (frosting). Figure 11 shows a micro-pitted carburized tooth. Figure 12 shows the surface and micro-pitting enlarged 430 times. This type of pitting is considered to be normal. SPALLING Spalling is a term used to describe a large or massive area where surface material has broken away from the tooth. In through-hardened and softer material, it appears to be a massing of many overlapping or interconnected large pits in one locality. See Figure 13. In surface-hardened material it manifests itself as the loss of a single or several large areas of material. The visual pit-like attributes are not observed (see Figure 14). Frequently the bottom of the spall appears to run along the case-core interface.

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FIGURE 11 Micro-pitting (frosting).

FIGURE 12 Micro-pitting (enlarged photograph). Spalling is caused by high contact stresses, possibly associated with proud areas of the tooth surface. With surface-hardened gear teeth, surface or subsurface defects or excessive internal stresses from improper heat treatment also can cause spalling. Case crushing is another form of spalling associated with heavily loaded case-hardened gears. It appears as long longitudinal cracks on the tooth surface which may subsequently break away. It often occurs suddenly, without warning signs, on only one or two teeth of the pinion or gear. The cracks differ from those of pits in that they not only extend below the hard case, but most of its depth is in the softer core material. The cracks in the case generally are perpendicular to the surface. Figure 15 shows an example of this failure mode. Failure may be due to insufficient case depth, insufficient core hardness, high residual stresses, or too high loading.

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FIGURE 13 Spalling, through-hardened gear.

FIGURE 14 Spalling, surface-hardened gear. Worm gear endurance tests have been run to approximately 20,000 hours. While varying degrees of surface destruction may be severe when compared with that of a helical gear the worm gears nevertheless survived the test. From this experience we can conclude that worm wheels incurring this amount of destruction can still perform satisfactorily. Figure 16 illustrates this surface deterioration.

WEAR “Wear” is a general term describing loss of material from the contacting surface of a gear There are varying degrees of wear, which can be measured in terms of thousandths of an inch, per million or 10 million contact cycles, ranging from light to moderate to excessive wear.

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FIGURE 15 Case crushing.

FIGURE 16 Worm gear pitting. DEGREES OF WEAR Polishing or light wear: Figure 17 shows the slow loss of metal at a rate that will little affect satisfactory performance within the life of the gears. It is a normal, very slow wear-in process in which asperities of the contacting surfaces are gradually worn until very fine, smooth, conforming surfaces develop. Polishing or light wear can occur by either abrasive or adhesive mechanisms when thin oil films or boundary lubrication conditions prevail, usually on slow speed applications. Moderate wear, sometimes called normal wear, progresses at a rate slow enough that it will little affect satisfactory performance of the gears within their expected life. Tooth contact patterns indicate that metal has been removed from the entire tooth surface, but generally more from the dedendum areas. The operating pitch line begins to show as an unbroken line. Surface-hardened gears, because of their high surface hardness, manifest less wear than do through-hardened gears. The judgment of moderate wear is somewhat subjective because performance requirements and expectations vary over the broad spectrum of industrial gear drive applications. Figure 18 illustrates wear observed on gears early in their operational history. Subsequently, these gears continued to give years of additional satisfactory service. Copyright © 1997 CRC Press, LLC.

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FIGURE 17 Initial pitting and light wear.

FIGURE 18A (1 YEAR) Typical tooth surface appearance. Moderate or normal wear depends on the lubrication regime, nature of the load, surface hardness and roughness, and on contaminants present in the lubricating oil which might promote abrasive or corrosive wear. The limitations of usable oil viscosities and the speed requirements of the application often dictate that the gears must operate in the boundary lubrication regime where metal-to-metal contact occurs and wear takes place. Copyright © 1997 CRC Press, LLC.

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FIGURE 18B (1½ YEARS) (Continued).

FIGURE 18C (4 YEARS) (Continued). Moderate wear is normal and usually no remedial action is required other than normal maintenance of the lube system. If contaminants are present, filtering or more frequent lube changes are appropriate. Excessive or destructive wear (see Figure 19) is surface destruction that has changed the tooth shape to such an extent that smoothness of meshing action is impaired and life is appreciably shortened. Continued operation results in still greater wear and may eventually lead to tooth breakage. The occurrence of such wear early in the operational history can be caused by excessive loads, contaminated oil, or too light an oil viscosity. Excessive wear incurred over a long period of operational history would be considered an advancement of normal wear from the moderate to the excessive degree and may not be detrimental to the operation of the gear drive.

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FIGURE 19 Excessive wear.

FIGURE 20 Abrasive wear, through-hardened gear. TYPES OF WEAR Abrasive wear, sometimes called cutting wear, occurs when hard part icles slide and roll under pressure across the tooth surface. Hard particle sources are: dirt in the housing, sand or scale from

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castings, metal wear particles from gear teeth or bearings, particles introduced into housing when filling with lube oil, and particles infiltrating into unit in service. Figure 20 shows abrasive wear of a through-hardened gear caused by massive loss of hard surface material from tapered bearing surfaces. Gear teeth surfaces hardened after cutting sometimes have a rough surface that may wear softer mating teeth. Figure 21 shows through-hardened teeth that were worn away in just a few hours from the flame-hardened teeth which had rough surfaces due to a sand-blast cleaning operation.

FIGURE 21 Abrasive wear, through-hardened pinion. Through-hardened pinions (350–390 BHN) run against flame-hardened gears.

FIGURE 22 Scratching. Scratching is a form of abrasive wear, characterized by short scratch-like lines in the direction of sliding. This type of damage is usually light and can be stopped by removing the contaminants that caused it. See Figure 22. Copyright © 1997 CRC Press, LLC.

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Adhesive wear results from high attractive forces of the atoms composing each of two contacting, sliding surfaces. Teeth contact at random asperities, and a strong bond is formed. The junction area grows until a particle is transferred across the contact interface. In subsequent encounters, the transferred fragment fractures or fatigues away, forming a wear particle. Adhesive wear depends upon the bond strength, which relates to the physical chemistry of the contact material and lubricant, on the load, and on the material hardness. Figure 23 shows typical surfaces of a surface-hardened gear that has undergone adhesive wear.

FIGURE 23 Adhesive wear, surface-hardened gear.

FIGURE 24 Scoring. Scuffing wear is adhesive-type wear occurring at normal temperatures, where smooth burnished appearing radial striations are observed in the direction of siding on the tooth surfaces. The texture of lower hardness through-hardened teeth is more coarse than that of higher hardness throughhardened or surface-hardened teeth of Figure 23. It can appear where tooth pressures are high and oil films are in the boundary regime and where speeds are slow enough that high-contact temperatures do not occur. This type wear can be reduced by increased oil viscosities where applicable or by reduced load. Scoring, Figure 24, is the smearing and rapid removal of material from the tooth surface resulting from the tearing out of small particles that become welded together as a result of oil film failure and high temperature metal-to-metal contact in the tooth mesh zone. After welding occurs, sliding forces tear the metal from the surface, producing a minute cavity in one surface and a projection on the other. The wear initiates microscopically; however, it progresses rapidly. Scoring is sometimes referred to as galling, seizing, or scuffing. Copyright © 1997 CRC Press, LLC.

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Scoring most frequently occurs in localized areas on the tooth where high contact pressure exists or at the tip or root where sliding velocities, and hence contact temperatures, are high. This mode of wear is usually associated with high pitch line velocities and is not common in the lower hardness through-hardened gears running at normal commercial speeds. The direct causes of scoring are high contact temperature and pressure and marginal lubrication. Scoring can sometimes be prevented by use of more viscous oil or by an EP type oil. Localized high contact pressure can be relieved by improved finishing of tooth surfaces. Sometimes profile or face modifications are required on highly loaded teeth to minimize high localized pressures. Welding is a hybrid form of scoring and pitting, where pit cracks are formed on the gear member with pit bodies subsequently adhering or welding to the pinion member. The two then run together with the profile formed by the original involute and the resultant pits bodies and pit cavities (see Figure 25). This phenomenon is thought to occur at high-load, low-speed and at marginal lubrication conditions where high contact temperatures prevail, but classical scoring does not occur. Increased oil viscosities or EP lubricants may help. Reduced loads will aid. Care must be exercised that axial displacement of the mating pinion and gear does not occur, as high localized pressure can result from the mismatching of high and low profile spots which could cause fracture. If gearing is disassembled and reassembled, the tooth surfaces should be dressed to remove proud bumps. Wavy tooth wear is occasionally observed on gears. Teeth can be observed to have wavy or undulating surfaces either by light reflection or by profile and lead checks. The crests and valleys of the waves usually lie parallel to the inclined lines of helical contact. See Figure 26. This wear pattern is thought to be caused by vibratory loads occurring in the system in which the gears are operating.

FIGURE 25 Welding, metal transfer gear to pinion. Copyright © 1997 CRC Press, LLC.

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FIGURE 26 Wavy tooth wear.

FIGURE 27 Surface bumps. Surface bumps are occasionally experienced as shown in Figure 27. The cause of such phenomena has not yet been defined. As with welding, axial movement could cause localized high tooth pressures which could fracture teeth. The tooth surfaces should be dressed to remove proud bumps. Wear pads. Pinion elements of a gear set are frequently made slightly wider than the gear element. As wear occurs, unworn pads are left at the tooth ends of the pinion. These cause no problem as long as axial positioning of gears is properly maintained. If unit is disassembled for some reason and subsequently reassembled, these pads on the ends of the pinion should be ground flush to the worn surface to assure that heavy contact and possible tooth fracture do not occur from this source. Furrowing has been observed on the working tooth surfaces of coarse-textured, low DP, “as hobbed” mill pinion teeth shortly after being put into service. To the unaided eye, it appears as a lattice of fine hair lines oriented generally in the root-to-tip direction. They are aligned in rows across the face of the tooth in an order consistent with the hobbing texture pattern (see Figure 28). The cause of furrowing has not been definitely established. Under the microscope, furrows appear as round bottom channels looking as if they were formed by hydraulic erosion. At present there are no recognized problems emanating from furrowing. It is thought to be prevented by minimizing localized contact pressure and by maintaining adequate lube films. Copyright © 1997 CRC Press, LLC.

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FIGURE 28 Furrowing.

PLASTIC FLOW Plastic flow is cold working of tooth surfaces caused by high contact stresses and the rolling and sliding action of the mesh. It is a surface deformation resulting from the yielding of the surface and subsurface material. Usually associated with the softer gear materials, it can occur in heavily loaded case-hardened gears as well. Cold flow occurs when surface and subsurface material show evidence of metal flow. Often surface material has been worked over the tips and ends of the gear teeth, giving a finned appearance. See Figure 29. This is sometimes called rolling or wire edging. Sometimes the tooth tips are heavily rounded-over and a depression appears on the contact tooth surface. Peening, another form of plastic flow as shown in Figure 30, is caused by excessive loading due to impact loading.

FIGURE 29 Plastic flow, rolling. Under heavy load, the rolling and peening action of the mesh cold works the surface and subsurface material. The sliding action tends to push or pull the material in the direction of sliding if the contact stresses are high enough. The dents and battered appearance of the surface are a result of dynamic loading due to operation while the profile is in the process of deteriorating from a combination of cold-working and wear. Failures of this type can be eliminated by reducing the contact stress and by increasing the hardness of the contacting surface and subsurface material. Increasing the accuracy of tooth-to-tooth spacing and reducing profile deviations will give better tooth action and reduce dynamic loads. If the high-contact stress is caused by mounting deflections or helix-angle error, these conditions should be corrected. Rippling is a periodic wave-like formation at right angles to the direction of sliding or motion (Figure 31). It has a fish-scale appearance and is usually observed on surface-hardened gear surfaces, although it can occur on softer tooth surfaces under certain conditions. Rippling is not always considered a surface failure, unless it has progressed to an advanced stage. Copyright © 1997 CRC Press, LLC.

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FIGURE 30 Plastic flow, peening.

FIGURE 31 Rippling. High contact stresses under cyclic operation tend to roll and knead the surface, causing the immediate surface material to-ripple. This type of failure is usually associated with slow speed operation and an inadequate oil film thickness. The combination of high contact stress, repeated cycles and an inadequate lubricating film will produce a rippled surface. Although rippling can be produced as a wear phenomenon, it most often is associated with a considerable amount of plastic flow. If the gear material is soft, rippling can be prevented by case-hardening the tooth surface. Also, reduction in contact stress will reduce the tendency of the surface to ripple. Since the lubricating film is marginal, an extreme-pressure additive in the oil and an increase in oil viscosity may be beneficial. An increase in rubbing speed is sometimes helpful. Ridging is the formation of deep ridges by either wear or plastic flow of surface and subsurface material (Figure 32). It shows definite peaks and valleys or ridges across the tooth surface in the direction of sliding. Ridging is caused by wear or plastic flow of surface and subsurface material due to high contact compressive stresses and low sliding velocities. It is often present on heavily loaded worm and wormgear drives and on hypoid pinion and gear drives. Ridging may occur on low-hardness materials and may also occur in high-hardness materials if the contact stresses are high, such as in case-hardened hypoid gear sets. Copyright © 1997 CRC Press, LLC.

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FIGURE 32 Ridging. Ridging can be prevented by reducing the contact stress, increasing the hardness of the material, and using a more viscous lubrication oil with extreme-pressure additives. In drives that do not have circulating systems, it is also helpful to change the oil often and to ensure that no foreign particles remain in the lubricant.

BREAKAGE Breakage is the ultimate gear failure. Bending loads on gear teeth usually cause the highest stresses at the root fillets and at the tooth profile/root fillet junctions. A gear tooth is a cantilever plate with tensile stresses on the contact side of the tooth and compressive stresses on the opposite side. If the tensile stresses at the critical location are allowed to exceed the endurance strength of the tooth material, fatigue cracks will eventually develop and with continued operation will ultimately progress to the point where the tooth will break away from the rim material. Classical tooth root fillet fatigue fracture is the most common fatigue breakage mode (Figure 33). The crack originates at the root fillet on the tensile side of the tooth and slowly progresses to complete fracture, either along or across the tooth. The faces of these fractures are usually characterized by a series of contour lines or “beach marks” caused by the progressing crack “front.” They indicate the position of the advancing crack front at a given time. As the section gradually weakens, the crack progresses further with each load cycle, and the beach marks become more coarse. The focal point of these marks often locates the origin of the fracture.

FIGURE 33 Tooth fracture (root fillet fatigue). Copyright © 1997 CRC Press, LLC.

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FIGURE 34 Tooth fracture (low cycle fatigue). Fatigue fractures result from repeated bending stress above the endurance limit of the material. If the tooth contact pattern appears even across the entire face, system overloads would be suspect. If the contact pattern is confined or “heavy” in the region of the fracture and at one end of the tooth, an alignment problem of the bearing would be suspect. When contact is good, system load must be reduced, or increased strength rating of the gearing is needed. If contacts indicate localized loading, gear alignment or face modifications should be examined. Low cycle fatigue or impact fractures are due to a low number of high load fatigue cycles or to a single very high load. Figure 34 illustrates such fractures. With lower hardness, more ductile materials, the fracture face is coarse, fibrous, and torn in appearance. With harder, less ductile material, the appearance may be smooth or silky. In some cases, a single overload may break out a tooth or several teeth. A more common occurrence is the plastic yielding of a group of teeth in one load zone from a high impact load. The plastic yielding displaces the pitch on this group of teeth with respect to the other teeth on the gears, thus subjecting them to abnormally high dynamic loads in subsequent operation. These teeth then develop very rapidly progressing fatigue cracks which soon lead to tooth breakage. This type failure is prevented by protecting the gearing from high impact or transient loadings. This may involve the use of controlled torque or resilient couplings in the connected drive train or it may require better control by the customer at the process being performed by the driven equipment. Pit-associated fractures occasionally originate in severely pitted areas, since pits can act as stress raisers and can be crack origins (Figure 35). Tooth tip chipping is a fracture mode where the top of a tooth will break away from the lower portion (see Figure 36). Failures of this kind may be caused by deficiencies in the gear tooth, which results in a high-stress concentration at a particular area. Sometimes flaws or minute grinding cracks will propagate under repeated stress cycling and a fracture will eventually develop. Foreign material passing through the gear mesh will also produce short-cycle failure of a small portion of a tooth. High residual stresses due to improper heat treatment can cause local fractures that do not originate in the tooth root section. It is difficult to prevent failures of this type except by good design and manufacturing practices. If trouble is encountered, the gear surfaces should be checked for possible previous damage that may have contributed to local stress risers. The history of heat treatment and manufacturing techniques should be reviewed to ensure that proper processing was carried out during all steps of the manufacturing cycle. Cleanliness of the gear material should also be examined. Worm wheel fractures are occasionally encountered. Figure 37 shows a tooth root-rim fracture on a bolt-on type worm wheel. Such failures are rare and are indicative of overloading.

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FIGURE 35 Fatigue fracture emanating from pit.

FIGURE 36 Tip chipping, through-hardened gear. Fracture due to failure of associated parts. Sometimes a severe load maldistribution of load on gear teeth can occur from damage to associated parts. Figure 38 illustrates a pinion run for a time after a severe bearing failure. Load shifted to one end of the teeth and they subsequently broke away. The process repeated itself two more times before the element was finally removed from service. Similar fractures could occur from a shaft that is severely bent or broken.

FAILURES ASSOCIATED WITH PROCESSING Quenching cracks can develop with some materials during heat treatment and after quenching of the gear blank in a quenching medium. Often these cracks are visible to the naked eye (Figure 39). They may run across the top land of the tooth or be radial in direction at the ends of the teeth. When these are found, the material and heat treat specifications should be reviewed and the processing procedures reexamined.

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FIGURE 37 Worm wheel tooth fracture.

FIGURE 38 Tooth fracture from bearing failure.

FIGURE 39 Quench cracks, through-hardened material.

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FIGURE 40 Grinding cracks, enlarged. Grinding cracks may form on the gear tooth surfaces due to process grinding, Figure 40. These surface cracks are usually in a definite pattern or network and often have the appearance of a series of short cracks laying parallel to each other. Grinding cracks may also have the appearance of a chicken-wire mesh. The cause may be excessive grinding pressures or may be a metallurgical structure which is prone to cracking. Sometimes grinding cracks are latent and do not show up until after several hours of shelf life or after operation under load. When found, an examination of the processing procedures is in order.

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and Particle Count 85 Filtration Classifications Charles A. Moyer CONTENTS Particle Size Distribution.........................................................................................................................1009 Filter Selection............................................................................................................................................1011 References...................................................................................................................................................1016

To provide longer life and increased performance, it is imperative to start equipment with as clean a fluid system as possible and to maintain the fluid cleanliness at a level commensurate with the costs of operating the equipment or the overall process. Whether the fluid is air, various fuels, or lubricants, filters are the paramount tool to maintain a clean system. Filters may also remove contaminants that are left on machine elements, are formed within various contacts, or enter the system from the environment.

PARTICLE SIZE DISTRIBUTION Contaminants are any matter that reduces system performance. They can be acids, water, oxidized lubricants, waxes, asphaltenes, bacteria, manufacturing residue, and various debris. Debris can be sand, metal turnings, grinding residue, wear particles, i.e., hard contaminants that can cause wear, fatigue, or failure to the working surfaces within operating equipment. The particles found in many systems range in size from over 200 µm (0.008 inch) to less than 5 µm (200 microinch [µin.]). The number of particles from small to large size tend toward an exponential distribution for many typical debris samples taken from operating fluid systems.1 Figure 1 is a simple illustration of five different debris distributions plotted on special paper based on ISO Standard 4406. Distribution No. 5 represents a milliliter sample of lubricant from a bearing life test machine lubricant sump, before test, but with AC fine test dust (debris) added and mixed in the lubricant. Nos. 1 and 3 are average debris particle distributions after a 40- and 3-µm filter, respectively, based on oil samples taken during the running with the contaminated oil. For comparison, distributions nos. 2 and 4 are from oil samples after a 40- and 3-µm filter during test runs with standard lubricant, that is, with no debris added to the lubricant sump. It is clear from these curves that filters can remove significant contaminant from a system. Distribution #2 (40-µm filter, but “clean” oil) and distributions #3 and #4 (3-µm filters) had much less contaminant present. Even distribution #1 (40-µm filter) had fewer larger particles than the dirty oil in distribution #5. The center numbers in Figure 1 refer to the ISO 4406 hydraulic cleanliness rating system based on both the number of particles in a distribution larger than 5 µm and larger than 15 µm in a 1-mm fluid sample. Table 1 shows the ISO 4406 particles range from 0.01 to 160,000 in 24 steps, and each step is a range or rating number. In Figure 1, distribution #1 has 5 µm = 25 and 15 µm = 18, Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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FIGURE 1 Five examples of debris distributions.1 or a cleanliness rating of 25/18 while distribution #2 has a rating of 13/10. From the distributions and cleanliness ratings alone, it is clear that #3 is better than #1. For these particular tests, debris did not influence the bearing failures, so 40 µm was adequate for the “system,” but for smaller bearings the 3-µm filter may have been required. Table 2 gives a similar listing of particulate contamination levels per SAE 749D, which has also been used in the past by various machinery builders and operators.

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FILTER SELECTION Primary attributes of the right filter for an application should be its removal efficiency, its contaminant-holding capacity, pressure drop across the filter, and its structural integrity.3 Reputable filter manufacturers can provide the above characteristics of their various filters to help in the selection of the appropriate one for your system. Laboratory tests are made with one pass or multiple-pass (recirculation) modes. Since data from these tests do not directly match actual field performance, a relationship between laboratory and field results must be gained through experience, somewhat as shown on Figure 1. Details of testing are given in Reference 3. Figure 23 shows the size distributions of four standard contaminants used in air filter evaluations by the three countries listed. AC fine test dust (ACFTD) and coarse dust have been used in various other tests to evaluate filters. As to airborne debris or dust concentrations, Figure 3 gives typical ranges for service or working environments.3,4 In the same vein for hydraulic systems, contaminants coming from wear, ingested from outside the system, or generated in the equipment manufacture can cause severe loss of performance, restrict fluid flow, and lead to component failure.3 Copyright © 1997 CRC Press, LLC.

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Figure 43,5 shows both the ISO 4406 contaminant code numbers based on either 5 or 15 µm particle size or larger and particle distributions based on cumulative number of particles greater than a particular size. The curve marked 1.0 means a sample distribution having 1.0 mg/1 ACFTD and the curve marked 100 means a distribution having 100 mg/1 ACFTD. In actual applications, the contaminant shapes and densities may differ and deviations from the ACFTD curves can be expected. The parallel curves in this figure were developed by Oklahoma State University Fluid Power Research Center (FPRC).6 The Fluid Power Research Center has run a series of tests to determine acceptable cleanliness levels for hydraulic pumps,7 and these results are the dashed lines on Figure 4. Filters need to be selected with high efficiency to keep within the cleanliness limits to avoid wear with different pumps. Like Figure 4, Figure 5 has the curves developed by FPRC6 for various working environments. Efficiency and performance of various filters can be compared with Figure 6.3,5 Efficiency is determined from the Beta rating system based on a ratio of the number of particles (Nu) larger than x(u) micrometers upstream of a filter to the number of particles (Nu) larger than x(d) micrometers downstream of the filter. That is,

For example, for 500 particles upstream larger than 10 µm (before the filter) and for 25 particles downstream (after the filter) larger than 10 µm, the Beta ratio is:

The efficiency is [(Beta - 1)/Beta] times 100%:

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FIGURE 2 Particle size distributions of standard contaminants. (From Jaroszczyk, T., Verdegan, B., and McBroom, K., in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. With permission.)

FIGURE 3 Environmental classification of air in various environments. (From Jaroszczyk, T, Verdegan, B., and McBroom, K., in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. With permission.) If a filter is required to remove 95% of the particles larger than 10 µm, then from Figure 6 the media (typical materials in hydraulic filters5) that will meet the need are 1, 3, and 5. Since field

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FIGURE 4 Methods of characterizing hydraulic oil contaminant levels. (From Jaroszczyk, T., Verdegan, B., and McBroom, K., in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. With permission.) performance is usually lower than in laboratory tests, filters should be selected with some safety factor.3 For detailed information covering types and designs of filtration units, the handbook Filtration, Principles and Practices8 would be a good starting point. Additional references are also listed that cover filtration (References 9 through 14). Reference 14 proposes an addition to ISO 4406 rating system. Besides particle sizes larger than 5 and 15 µm, they would include 2 µm (ISO 4406 number range 16). According to Pall Industrial Hydraulics Corporation, this third particle size will help separate fluids with and without high silt content in the particle range 0 to 3 µm.2,14 References 15 through 19 cover debris, particles, and particle size distributions, and how they impact bearings and other mechanical systems.

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FIGURE 5 Contaminant levels in hydraulic oil. (From Jaroszczyk, T., Verdegan, B., and McBroom, K., in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. With permission.)

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FIGURE 6 Performance curves for hydraulic filters. (From Jaroszczyk, T., Verdegan, B., and McBroom, K., in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. With permission.)

REFERENCES 1. Moyer, C. A., The influence of debris on rolling bearing performance: identifying the relevant factors, SAE Trans., 96(3), 870–879, 1987–1988. 2. Arndt, B., Too many rating systems confuse filtration needs, Mach. Des., March 12, 1992. 3. Jaroszczyk, T., Verdegan, B., and McBroom, K., Cartridge filtration, in Filtration, Principles and Practices, 2nd ed., Matteson, M. J. and Orr, C., Eds., Marcel Dekker, New York, 1987, chap. 10. 4. Cockle, G. R., Houser, W. N., and Koch, E. M., Eds., Car and Locomotive Cyclopedia of American Practices, 3rd ed., Simmons-Boardman, New York, 1974. 5. McBroom, K., Upgrading hydraulic system filtration, Plant Eng., June 28, 1984. 6. Fitch, E. C, Encyclopedia of Fluid Contamination Control, Technical Communications Stillwater OK, 1978. 7. Fitch, E. C. and Hong, I. T., Pump Contaminant Sensitivity — an FPRC Position Report, Fluid Power Research Center, Stillwater, OK, 1984. 8. Matteson, M. J. and Orr, C., Eds., Filtration, Principles and Practices, 2nd ed., Marcel Dekker New York, 1987. 9. Kolp, R., Comparing the effectiveness of safety screens, Mach. Des., September 10, 1992. 10. Staley, D., Correlating Lube Oil Filtration Efficiencies with Engine Wear, SAE 881825, Truck and Bus Mtg. and Exposition, Indianapolis, IN, November, 1988. 11. Needelman, W. and Zaretsky, E., New equations show oil filtration effect on bearing life, Power Trans. Des., August, 1991.

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12. Pohl, M., Selecting a particle size analyzer: factors to consider, Powder Bulk Eng., February, 1990. 13. Verdegan, B. M., Jaroszczyk, T., and Stinson, J. A., Interpretation of filter ratings for lubrication systems, STLE Lubr. Eng., 44(5), 424–430, 1988. 14. Needleman, W. N., Filtration, CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 71–87. 15. Sayles, R. S., Debris and roughness in machine element contacts: some current and future engineering implications, Proc. Inst. Mech. Eng. J. Eng. Tribol., 209(J3), 149–172, 1995. 16. Dawson, D., Taylor, C. M., Childs, T. H. C, Godet, M., and Dalmaz, G., Wear particles: from the cradle of the grave, in Proc. 18th Leeds-Lyon Symp. Tribol, Elsevier, London, 1992. 17. Dwyer-Joyce, R. S., Hamer, J. C, Sayles, R. S., and Ioannides, E., Surface damage effects caused by debris in rolling bearing lubricants, with an emphasis on friable materials, Inst. Mech. Eng. Conf. Proc. Rolling Element Bearings — Toward 21st Century, p. 1–8, Mechanical Engineering Publications, London, 1990. 18. Beerbower, A., Wear rate prognosis through particle size distribution, ASLE Trans., 24(3), 288– 292, 1981. 19. Fitzsimmons, B. and Clevenger, H. D., Contaminated lubricants and tapered roller bearing wear, ASLE Trans., 20(2), 97–107, 1977.

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86 Life of Oils and Greases E. Richard Booser CONTENTS Industrial Oil Life....................................................................................................................................1018 Grease Life................................................................................................................................................1022 References.................................................................................................................................................1027

Service life becomes a significant factor in selecting and applying oils and greases. This life factor not only sets relubrication schedules, but also influences cooling considerations, the feed system, and lubrication quantity in design and operation of equipment. Lubricant life predictions involve large uncertainties because of the many variables in lubricant composition, contamination, and machine operating details. While appreciating these uncertainties, some guidelines follow for estimating lubricant life: first for industrial oils, then for greases in ball and roller bearings. Automotive and related reciprocating engine applications involve special considerations and are covered in Chapters 79-81.

INDUSTRIAL OIL LIFE Lubricating oils circulating in industrial units such as turbines, compressors, electric motors, and hydraulic systems deteriorate primarily by oxidation. This aging process, in which oil chemically breaks down by reaction with atmospheric oxygen, undesirably brings increased oil acidity and viscosity, darkening color, and surface deposits.1.2 Peroxide formation is the first step in oxidation of either petroleum or most synthetic oils involving primarily hydrocarbon structures. This initiates a free radical chain mechanism which leads to oxygen-containing molecules such as hydroperoxides, aldehydes, ketones, alcohols, esters, and acids. Oxidation-inhibiting additives are employed to control this oxidation, either by attacking hydroperoxides formed in the initial oxidation step or by breaking the chain reaction mechanism. Aromatic amines, hindered phenols, and alkyl sulfides are oxidation inhibitors that function by one of these mechanisms. Metal deactivators provide a third type of oxidation control: they keep metal surfaces and soluble metal salts both from catalyzing oil oxidation and also from polymerizing oil oxidation products to produce sludge and varnish. Combination of phenolic and aromatic amine inhibitors with a metal deactivator gives superior life for demanding industrial applications. As a simplified view, useful life of most industrial oils continues through an induction period during which the oxidation inhibitor is slowly consumed by oxidation, evaporation, or other physical and chemical effects. While subsequent additions of new oil or oxidation inhibitor will delay the process, the end of useful life is reached when the oxidation inhibitor is finally exhausted and oxidation reactions accelerate. Even without an added oxidation inhibitor, an induction period is commonly provided with industrial mineral oils from the inhibiting effect of naturally occurring Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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sulfur and polynuclear aromatic hydrocarbon components. This induction period is shortened by higher temperature, more oxygen availability, catalytic metals, and water. Standardized laboratory tests used in the U.S. for evaluating oil oxidation life include ASTM D943, D2272, D2893, and D4742; and Federal Test Method Standard 791, Method 5308.6. Other test procedures using various combinations of temperature, oxygen, and catalysts include the ALCOR deposition test, International Harvester BT10, IP280, IP331B, and the U.S. Army Mobility Equipment R and D Method.3 OXIDATION LIFE OF MINERAL OILS Based on test results, oxidation life L (h) drops by a factor of 2 for a 10°C temperature rise in the 100 to 150°C range. Using this factor gives k2 = 4750 in the following Arrhenius equation for chemical reaction rates:

Table 1 gives typical values of k1, for industrial lubricants subjected to agitation by air under laboratory conditions without the presence of catalysts or contaminants. Lubricant types included in Table 1 are (1) uninhibited oils such as used in a once-through system; (2) EP gear oils; (3) conventional hydraulic oils used in many industrial systems; (4) premium rust and oxidation-inhibited turbine oils employed for long life in turbines, compressors, and electric motors; and (5) the longest service life oils available following refining by severe solvent extraction or hydrocrack-ing. Temperatures calculated from Equation 1 for several oxidation life periods are also given in Table 1.

Variations between similar types of products from different suppliers vary so widely that Table 1 gives only a general indication of oxidation life to be expected. Individual suppliers should be contacted for detailed performance experience with their products. LIFE REDUCTION FACTORS Unfortunately, general correlation of laboratory oxidation test results with field experience has yet to be realized. Lubrication service usually involves factors which reduce the life values of Table 1. The combined effect of water plus copper and iron catalytic surfaces, for instance, drops life by a factor of about 6 as reflected in the ASTM D943 Turbine Oil Stability Test. Experience suggests a similar reduction factor of about 2 to 5 for steam turbine service, and about 3 for electric motors. For land-based heavy-duty gas turbines a factor of 10 is more appropriate to cover added degradation from hot spots, oxidation inhibitor evaporation, and other deterioration effects.

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With service life sometimes only being one tenth that for uncatalyzed oxidation, oil life in Table 1 should be dropped by a factor of 10 to reflect field service conditions (or the limiting temperatures in Table 1 should be dropped about 33°C using a life factor of 2 per 10°C). Usually the temperature of the reservoir, where about 75% of the system oil charge C is commonly located during operation, can be used to characterize oil oxidation life L to be expected in a system. For considering the influence of each machine element n, its individual temperature can be considered to generate deterioration rate 1/Ln for average volume of oil Cn being held at n.

Consider, for instance, an individual bearing assembly with 20 gallons of turbine oil held at 138°C (280°F). With k1 = -8.45 from Table 1 giving oxidation life L of 1222 h. from Equation 1, its deterioration factor C/L = 20/1280 = 0.016; while 2000 gal in the main reservoir at 71°C (160°F) with L = 228,000 h gives a C/L deterioration factor of only 0.009. Thus, nearly twice as much oil deterioration by loss of oxidation inhibition (0.016/0.009) would be expected in the 20 gal in the hot bearing assembly as in the 2000 gal held at a lower temperature in the reservoir. Evaporation of a low molecular weight phenolic oxidation inhibitor such as ditertiarybutyl paracresol (DBPC) at the hot bearing would make its deterioration effect even greater. MONITORING REMAINING USEFUL OIL LIFE As antioxidant additives are depleted with equipment operating time, they eventually become ineffective and allow major increases in oil acidity (TAN), oil viscosity, color, and varnish deposits. Length of time from lubricant sampling until large property changes occur, the remaining useful lubricant life, can be evaluated by various techniques.5-7 Fourier transform infrared spectroscopy (FTIR) is the most widely used instrumental technique for measuring remaining oxidation inhibitor concentration. The amount of light absorbed in the 2-to 50-micron (µm) wavelength range is used to identify the type and concentration of each type of inhibitor.8 With two different types of oxidation inhibitors in antiwear hydraulic fluid undergoing different depletion rates in Figure 1, the drain period can be extended more efficiently by selective replacement of the inhibitor type undergoing accelerated depletion.5 The ASTM D2272 rotating bomb oxidation test (RBOT) is used extensively to monitor the remaining antioxidant life of turbine oils using phenolic and amine type inhibitors. While a RBOT induction period of 200 min has been used to qualify candidate oils for steam turbine service, induction times range up to 2800 min for some fresh gas turbine oils.3 Generally, a drop below 50 min in the RBOT test or a drop of half or more in oxidation inhibitor concentration would call for replenishment of the inhibitor to its concentration which existed in the fresh oil charge. For routine evaluation of remaining oil life, more rapid electrochemical methods, microscale oxidation tests, differential thermal analysis, and high-pressure differential scanning calorimetry have been used.5 Monitoring is largely ineffective when based on measurement of viscosity, acidity, color, dielectric strength, and particulates. These values commonly change significantly only after protection by the oxidation inhibitor has been lost and bulk oil degradation has begun. SYNTHETIC OIL LIFE Figure 2 compares oxidation life for inhibited synthetic oils of various types in comparison with inhibited mineral oil in an air environment in contact with steel.9 While this comparison indicates longer life for all synthetics, their potential superiority over mineral oils is lost under some service conditions. Degradation by hydrolysis, for instance, will lead to shorter life for some phosphates, silicates, and esters, even when exposed simply to atmospheric humidity. Where their unique temperature range or fire resistance is required, special precautions may be necessary

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FIGURE 1 Percent remaining additive, viscosity (40°C), and total acid number (TAN) vs. stressing time at 150°C for a typical antiwear hydraulic fluid. (From Kauffman, R. E., in CRC Handbook of Lubrication and Tnbology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 98. With permission)

FIGURE 2 Life expectancy of inhibited lubricants in air. (From Beerbower, A., in STLE Spec. Publ., No SP-15, 1982, 58–69. With permission.) with appropriate additives, desiccants for the air, and absorptive filtration of hydrolysis products to minimize hydrolytic reactions. Copyright © 1997 CRC Press, LLC.

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GREASE LIFE Life of greases in rolling element bearings has been approximated both from observations in field service and from laboratory ball bearing tests such as ASTM D3336, the ROF of SKF, and the German FE9 covered by DIN 51821.10-18 Since life scatter ranges up to 10-fold even under seemingly identical operating conditions, calculation of expected grease life and setting of regreasing schedules must be approached on a statistical basis similar to that used for variability in ball and roller bearing fatigue life. MEASURES OF GREASE DETERIORATION The primary function of grease in a rolling bearing is to serve as a reservoir for lubricant to coat bearing surfaces with an EHD film thick enough to cover their surface roughness. Observations from widely different laboratory tests and from industrial, appliance, railroad, and aerospace applications have indicated that a grease commonly fails to fill this lubrication demand when half of its initial oil content is lost.10,15,16 Loss of half of the initial oil content corresponds to the following percentage soap (thickener) Sf in the grease at failure:

where So is the per cent thickener in the fresh grease. For a fresh grease containing 10% soap, failure would then be expected at Sf = 18%; for 25% initial soap content, Sf = 40%. A similar soap content at failure results from the 40% oil bleeding deterioration limit of Table 2, which gives Sf= 100 So/0.6. This drying and hardening of the grease, which leads to excessive noise, high friction, and increasing temperature rise, appears to involve oil loss through a combination of creepage, oxidation, and evaporation.10

Accelerated oil loss and deterioration in other physical properties follow the depletion of antioxidant, much as has been observed with lubricating oils themselves.15 Both acidity and molecular weight were seen to increase rapidly after antioxidant depletion. Bearing wear gave a rise in iron content in the drying grease with the approach of grease failure. Table 2 summarizes a number of methods for evaluating grease deterioration and limits where failure of bearings is to be expected. Quantity of grease applied in a bearing has a quiet variable effect on grease life. Komatsuzaki16 found that life increased 10-fold and was proportional to increasing grease fill in a 100-mm bore cylindrical roller bearing running at 3000 rpm and 100°C. While somewhat shorter life is available Copyright © 1997 CRC Press, LLC.

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with the limited grease volume in double-sealed and double-shielded ball bearings, only minor variation in life is experienced with changing the amount of the channeling-type greases commonly used. Kleinlein17 reported longest life with polyurea grease when extra grease was applied in space outside a bearing seal. GREASE LIFE-TEMPERATURE-SPEED RELATIONS Grease life in service can be related to temperature in the same general pattern given by Equation 1 for oil life. Under ideal operating conditions with light load and low surface speed, the log of 10% grease life L (h) is generally a function of the grease stability term A and the reciprocal of absolute temperature (273 + °C) in the first two terms on the right side of the following equation:10,11

With life rising typically by a factor of 1.5 per 10°C lower temperature in the 100°C range, B = 2450. There is some indication that this temperature factor rises to about 2/10°C for the range above about 150°C, with oxidation becoming the life-controlling factor. From test lives with premium greases formulated with mineral and a number of synthetic oils and various thickeners such as lithium hydroxystearate, complex metal soaps, and polyureas, A becomes -2.30.11 These A and B values are used in the life curves of Figure 3. Reduction in grease life with increasing surface speeds is reflected in the kfDN term (D mm bore times N rpm) on the right side of Equation 4. Velocity coefficients kf in Table 3 reflect relative lubricant needs of various bearing types with increasing speed.13 The higher values for a given bearing type apply to the larger cross-section series (higher load capacity) and the smaller values to lighter series bearings of lower load capacity. While this velocity term reflects typical experience for premium greases in open ball and roller bearings in electric motors and related equipment, the pattern for decreasing life with increasing DN values is quite variable with different greases. Results more sensitive to DN reported by bearing suppliers13,14 likely reflect behavior with greases at least partially selected both to accommodate close confinement with the ball complement in double-sealed bearings and also for adaptability to high temperatures. An informative analysis of dependence of grease life on bearing temperature in over 2000 fractional horsepower motors (bearings of 3.2- to 19-mm bore at speeds to 25,000 rpm) by Smith and Wilson18 gave the four distinctive regimes in Figure 4. For the “warm” temperature range from 50 to 200°C which covered most of the cases, the simplest correlation of mean life in hours was given by the following equation with a form similar to Equation 3:

Adding a speed factor of -4.32(10-6)DN on the right side of the equation, while reducing the A value by 15%, gave only a modest improvement in correlating the life results. Segregating results by the grease used, primarily two diesters and one silicone grease, also had relatively minor influence on the correlations. The four bearing temperature ranges in Figure 4 were characterized as follows: Hot: Above 200°C. Rapid oxidation led to shorter life with a log life slope with temperature twice as steep as in the warm range. Warm: Between 50 and 200cC. Most common operation. Cool: Between 50 and -10°C. Life dropped rapidly as temperature went down. Cold: Below -10°C. Life not influenced by temperature. The surprising decrease in life as temperature dropped below 50°C suggests that the lubrication effectiveness of the grease dropped as grease mobility diminished. This raises a question as to the Copyright © 1997 CRC Press, LLC.

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FIGURE 3 Ball bearing grease life vs. bearing temperature and kfDN speed factor. extent to which high-temperature laboratory grease life data can properly be extrapolated (as in Figure 3) to estimate much higher life at the lower bearing temperatures in industrial, transportation, and other nonmilitary equipment. With development of greases having increasing upper temperature limits, grease life is now often evaluated in laboratory tests at 125 to 170°C and higher to obtain reasonable test time. In recent tests,19 a thermally stable ester oil-polyurea grease gave 1000 hours life at 200°C, a 10fold life increase above that given by Equation 3. With ultimate oxidation stability afforded by perfluoroalkylpolyether greases, 3232 h life was provided at 176°C (350°F) and over 750 h at 260°C (500°F) in ASTM D3336 tests.20 Copyright © 1997 CRC Press, LLC.

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FIGURE 4 Grease-lubricated bearing life as a function of bearing operating temperature for fractional horsepower motors. (From Smith, R. L. and Wilson, D. S., L ubr. E ng., 36, 411–416. With permission.) Copyright © 1997 CRC Press, LLC.

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Unfortunately, the elevated test temperatures currently used in most grease life evaluations are above the temperature limit for many lithium and other soap-thickened greases which are commonly superior to high-temperature greases for long life and other performance characteristics in electric motors and related industrial applications with bearing temperatures below 100°C. LIFE VARIABILITY AND REGREASING Life L in Equation 3 represents a 10% likelihood of grease failure. With the usual log normal or Weibull grease life distributions experienced in tests, 1% failure requirement drops L in half, and 50% failures are encountered at twice L.10,17 Regreasing at half of 10% life L will normally provide insurance against grease drying or hardening to a degree where surface damage is encountered in a rolling bearing. The grease quantity G (g) suggested for use in relubrication is:14

where D = bearing outside diameter, mm; B = total bearing width, mm. For increasing demands from poor environmental conditions, vibration, shock, and high loads, grease life reduction factors in Table 4 give an adjusted 10% life La from the following:13

where the unadjusted 10% life L is obtained from Equation 3 or Figure 3.

LIMITATIONS AND GENERALIZATIONS Considering the frequently poor reproducibility of grease tests, the variability in rheology of greases and their batch-to-batch properties, and intangible differences in application environments, relations such as those outlined here for grease life calculations should be used only as guidelines.

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Their most reliable use is likely in extrapolations from established operating conditions. If bearing temperature drops from 80 to 60°C, for instance, the temperature term in Equation 3 indicates that 60% increase in life is to be expected. Regreasing schedules might be adjusted from Equations 3 and 6 for changes in bearing speed, size, and load. Considering use of a new grease might involve comparing its test life with that of a presently used grease based on data from bearing or grease manufacturers. Laboratory data should be used with caution, however, since they are obtained in relatively small ball bearings at speeds and temperatures above those of interest in most industrial applications. A grease showing promise from an initial check should then be subjected to a plant trial under the most severe conditions to be expected while observing factors such as noise, leakage, rate of drying and hardening, water and humidity effects, and initial temperature rise with overgreasing. Grease life estimates should be used only within the useable range of the grease and the bearing. Some lithium soaps, for instance, undergo a phase change at about 112°C and provide much shorter life at higher temperatures. Conventional radial deep-groove ball bearings and cylindrical roller bearings commonly have a DN speed limit (mm bore-rpm) with grease in the 250,000 to 300,000 range, above which very short life can be expected. At higher speeds, centrifugal action appears to throw the grease from the inner ring and cage surfaces so as to leave a thinner grease film than needed for EHD lubrication. This DN limit can be raised by use of precision bearings with their finer surface finish, fabric-filled plastic cages with their retained oil supply, or some channeling type greases.

REFERENCES 1. Klaus, E. E. and Tewksbury, E. J., Liquid lubricants, in CRC Handbook of Lubrication, Vol. 2, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1984, 229–254. 2. Hsu, S. M., Ku, C. S., and Pei, P. T., Oxidative Degradation Mechanisms of Lubricants, ASTM STP 916, American Society for Testing and Materials, Philadelphia, 1986, 27–48. 3. Smith, A. N., Turbine Lubricant Oxidation: Testing, Experience, and Prediction, ASTM STP 916, 1986, 1–26. 4. Lansdown, A. R., Selection of lubricants, in Industrial Tribology, Jones, M. H. and Scott, D., Eds., Elsevier, New York, 1983, 223–241. 5. Kauffman, R. E., Rapid determination of remaining useful lubricant life, in CRC Handbook of Lubrication and Tribology, Vol. 3, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1994, 89–100. 6. Fitch, J. C, Elements of an oil analysis program, CRC Tribology Data Handbook, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1997, chap. 78. 7. Saba, C. S., Gas turbine engine lubricant monitoring and analysis, CRC Tribology Data Handbook, Booser, E. R., Ed., CRC Press, Boca Raton, FL, 1997, chap. 78. 8. Coates, J. P. and Setti, L. C, Infrared Spectroscopy as a Tool for Monitoring Oil Degradation, ASTM STP 916, 1986, 57–78. 9. Beerbower, A., Environmental Capabilities of Liquid Lubricants, STLE Spec. Publ., No. SP-15, Society of Tribologists and Lubricating Engineers, Park Rodge, IL, 1982, 58–69. 10. Booser, E. R., Grease life forecast for ball bearings, Lubr. Eng., 30, 530–540, 1974. 11. Booser, E. R., When to grease ball bearings, Mach. Des., August 21, 1975, 70–73. 12. Grease life estimation in rolling bearings, Engineering Sciences Data Unit, London, ESDU 78032, 1978. 13. The Lubrication of Rolling Bearings, Pub. No. WL81 115/2 EC/ED, FAG Bearings Corp., Stratford, ON. 14. SKF Bearing Maintenance Handbook, Publ. 4100E, SKF Group, King of Prussia, PA, 1991. 15. Tomaru, M., Suzuki, T., Ito, H., and Suzuki, T., Grease Life Estimation and Grease Deterioration in Sealed Ball Bearings, NSK Tech. Paper 602204, Proc. JSLE Int. Tribology Conf., Tokyo, July 8 to 10, 1985. 16. Komatsuzaki, S. and Uematsu, T., Estimation of service life of grease in large size roller bearings, Lubr. Eng., 50, 25–29, 1994.

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17. Kleinlein, E., Grease test system for improved life of ball and roller bearings, Lubr. Eng., 48, 916–922, 1992. 18. Smith, R. L. and Wilson, D. S., Reliability of grease-packed ball bearings for fractional horsepower motors, Lubr. Eng., 36, 411–416, 1980. 19. Loderer, K., Lifetime Lubrication of Rolling Bearings under Extreme Conditions, National Lubricating Grease Institute, File No. 9524, Kansas City, 1995. 20. KrytoxR Oils and Greases, DuPont Bull. H-58505, 1996, Wilmington, DE.

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XI Toxicology, Environment, Safety and Health

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Environment, Safety and 87 Toxicology, Health Joseph M. Perez and Donald I. Hoke CONTENTS Introduction.....................................................................................................................................................1031 Trends in Toxicity of Lubricant Components.........................................................................................1032 Lubricants and Fluids in the Workplace....................................................................................................1033 Material Safety Data Sheet............................................................................................................................1034 Toxicology Testing..........................................................................................................................................1035 Environmental Exposure.............................................................................................................................1036 Waste Disposal................................................................................................................................................1037 Transportation.................................................................................................................................................1038 Other Applicable Laws and Regulations...................................................................................................1038 Biodegradability, Disposability, Recyclability, and Life Cycle Analysis...............................................1040 Conclusions......................................................................................................................................................1041 Appendix 1: Applicable Laws.......................................................................................................................1042 Appendix 2: Acronyms..................................................................................................................................1043 References.........................................................................................................................................................1043 Information Sources......................................................................................................................................1044

INTRODUCTION It is now recognized that a small percentage of chemicals can cause health and environmental effects. The law provided the initial push, but industry has now taken the initiative in controlling releases, developing hazard data, and communicating that information to users of chemicals. The Responsible Care program of the Chemical Manufacturers’ Association and Responsible Care associate programs of other trade organizations, including API, have provided impetus not only to comply with the pertinent laws but to go beyond that. In addition, the need for the practical engineer or scientist to work with the professional experts and to integrate health, safety, and environmental issues into their daily work is now the norm. It is important that these professionals have access to the proper information and knowledge about materials with which they work and the plethora of laws which may impact their work. Conservation, health, safety, and environmental pollution concerns have led to the creation of wide reaching legislation including, for example, the U.S. Congress Energy Policy and Conservation Act, Toxic Substances Control Act of 1976, Resource Conservation and Recovery Act of 1976, the Oil Recycling Act of 1980, and the subsequent implementation of many rules and regulations such as the OSHA Hazard Communication Rule. A list of the primary legislation covering this field is given in Appendix 1. This chapter is intended to provide a brief overview of health, safety, and environmental issues which may affect the users of all the various types of lubricants, hydraulic fluids, and metalworking Copyright © 1997 CRC Press, LLC. 0-8493-3904-9/97/$0.00+$.50

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fluids. These laws have potential impact in the workplace due to worker exposure and potential release of material into the surrounding air, during transportation to the workplace, when a material is spilled or released in any manner, and also when disposal of used material is being considered. Definitions for hazardous and toxic materials are critical and confusing since they vary from law to law. While some of the definitions are included here, a much more complete comparison of criteria may be found in Reference 1.

TRENDS IN TOXICITY OF LUBRICANT COMPONENTS In the past, concerns were expressed over the health effects of certain components in lubricants. Polynuclear aromatic hydrocarbons cause tumors in laboratory animals under some test conditions. Based on these findings, the International Agency for Research on Cancer (Volume 33) classified as carcinogenic any oil that was not severely solvent refined. The Occupational Safety and Health Administration (OSHA) in its Hazard Communication Rule mandated that oils which were not so refined must be classified and labeled as carcinogenic and later set guidelines for the classification of severely hydrotreated oils. Today, almost all companies use oils for manufacture of lubricants that are either severely solvent treated or severely hydrotreated. For a number of years, lubricants for automatic transmission fluids used an oil that contained a very high content of polynuclear aromatic hydrocarbons. The aromatic content kept the seal around the drive shaft pliable and induced sufficient swelling to prevent leakage. The use of this type of fluid was replaced in the 1970s by synthetic fluids. Other chemicals that raised concerns in the past included naphthylamine, tri-ortho-cresyl phosphate, and chlorinated naphthalenes. The major producers of lubricants discontinued most uses of these substances years ago. In the early 1980s, evidence of tumors and other effects were reported in laboratory animals exposed to nitrosamines and certain lower molecular weight chlorinated paraffins. Nitrosoamines can form by the reaction of sodium nitrite, a rust inhibitor used in metalworking fluids and coolants, and secondary amines which were used in many of these aqueous based systems. As a result, alternatives to sodium nitrite as rust inhibitors were developed and are now widely used. While the lower molecular weight chlorinated paraffins showed health effects, the higher molecular weight homologs did not. Despite this, the use of even the higher molecular weight homologs of this class of substances is declining. In the area of metalworking products where there is a high level of skin contact, precautions are still required for handling and use. Despite the move to safer lubricating oils and additives, prolonged contact with oils and other organic materials can lead to drying and cracking of the skin and the development of various types of dermatitis without proper handling and care. Where such contact is unavoidable, suitable protective skin creams or gloves can be used, if appropriate. Inhalation of the mists produced in some metalworking operations should be avoided. Inhaled mineral oils vapors can lead to an inflammation of the lungs or more severe problems after continued exposure. It should be noted that OSHA has established an exposure limit of 5 mg/1 of air for mineral oils and is considering a reduction to 0.5 mg/1 for metalworking fluids. Polychlorinated biphenyls (PCBs), because of their stability, were widely used in industrial heat exchangers, transformers, capacitors, and other electrical components. While PCBs were generally referred to as “transformer oils,” there was some use in special lubricant applications as fire resistant oils and also in some pipe-line turbines. To the authors’ knowledge, PCBs were not used as lubricant additive components because they did not offer any performance benefits. Concerns focused on the possibility that used oils and PCBs were not separated in some reclamation systems with the potential for ppm levels of contamination in lubricant base stocks. Disposal by low temperature burning of PCBs can result in formation of chlorinated dibenzodioxins (dioxin). Such an event was reported to have occurred in Germany in the mid-1980s. Both PCBs and dioxin are believed to cause adverse health effects. High temperature incineration of PCBs avoids this problem. The banning of PCBs in the U.S. was intended to address this problem.

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Essentially all lubricants today are mixtures of a base oil, usually petroleum based or a synthetic fluid derived from petroleum chemicals, and a combination of chemicals referred to as an additive or additive package used to meet specific performance requirements. The components of the additives normally have a very low order of toxicity either in mammalian species or aquatic species. Individual additive companies in their own testing programs have developed data that show the LD50 of most of these components to be greater than 1000 mg/kg. Even the most toxic additive components, in general, have LD50 values orders of magnitude above any level of concern. A paper presented in 1994 summarizes the known toxic effects of a number of important types of lubricant additive components.2 Data on the toxicity to selected fresh water organisms was presented at the November 1995 SETAC (Society for Environmental Toxicology and Chemistry) meeting in Vancouver, Canada and at the Esslingen Tribology Colloquium in January of 1996. Eye or skin irritation are of primary concern, but even these effects disappear at concentrations used in finished lubricant formulations. Skin sensitization is a concern with some additive components such as calcium sulfonates. It has been shown through human patch testing in a yet to be published work by one of the authors and his co-workers that it is low-base calcium sulfonates and not the high-base calcium sulfonates that are skin sensitizers. Similar to the irritation effect, this sensitizing effect generally disappears at the concentration used in finished lubricants. Although additive components are chosen on the basis of performance, increasing emphasis is being placed on their health effects, which has resulted in a significant increase in the amount of toxicity testing being conducted today. This will certainly lead to still safer lubricants and functional fluids.

LUBRICANTS AND FLUIDS IN THE WORKPLACE Since exposure to lubricants and other functional fluids is almost inevitable, it is important to understand potential health effects and proper protective measures to avoid excessive exposure. In general, lubricants and metalworking fluids have little toxic effect. The primary effects of concern are given in the previous section. The employer and workers and others should now look for information which is specific for the product being used. Despite the low order of health effects from these materials, there are laws and regulations which may apply. The foremost regulation for the workplace is the OSHA Hazard Communication Rule (29 CFR 1910). This rule has been judged to override state laws directed towards the workplace and requires that hazards be communicated by use of labels, material safety data sheets (MSDS), and worker training. Labels provide a worker with a first warning of potential hazards of any product being used in the workplace and must show the main hazards and toxic effects. More extensive information, especially with regard to toxic effects and precautionary and first aid measures are to be found on the MSDS. OSHA has ruled that the National Fire Protection Association diamond symbols may be used within a plant site, providing that workers have been trained to understand them. Although this Rule requires that an MSDS be provided for only those products which must be classified as hazardous, essentially all manufacturers provide an MSDS for each product or group of similar products even though they may not be classified as hazardous. The MSDS can be confusing and difficult to understand, so a brief description will be provided here as an initial guide. Anyone wishing to create an MSDS should refer to the guidelines developed by the Chemical Manufacturers Association and subsequently adopted as the ANSI standard for MSDSs. Copies of the standard may be obtained from the American National Standards Institute. This standard meets all the requirements of the OSHA Rule and provides for more information than does the OSHA Rule. In addition to the OSHA Hazard Communication Rule, care should be taken to observe other workplace requirements. In particular, exposures to chemicals require attention. OSHA has published a list of exposure limits for a variety of chemicals which are referred to as PELs or permissible exposure limits. It is also prudent to monitor the limits established by the ACGIH (American Conference of Government and Industrial Hygienists’ “Threshold Limit Values for Chemical Copyright © 1997 CRC Press, LLC.

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Substances and Physical Agents and Biological Exposure Indices”). These limits are referred to as TLV™s or threshold limit values. In general, both PELs and TLVs are established by applying safety factors of 10 to 1000, depending on whether the limits are based upon data from humans, laboratory animals, or in vitro tests.

MATERIAL SAFETY DATA SHEET Sale and shipment of commercial additive packages, lubricants, and functional fluids is often accompanied by the MSDS. The ANSI standard prescribes 16 sections and specific text for the headings, a format which has gained global acceptance through global chemical industry efforts and is the standard promoted by the International Labor Organization. The section headings and recommended order are: 1. Chemical Product and Company Identification 2. Composition and Information on Ingredients (normally just the ingredients contributing to the hazard of the product are reported) 3. Hazards Identification 4. First Aid Measures 5. Fire Fighting Measures 6. Accidental Release Measures 7. Handling and Storage 8. Exposure Controls/Personal Protection 9. Physical and Chemical Properties 10. Stability and Reactivity 11. Toxicological Information 12. Ecological Information 13. Disposal Considerations 14. Transport Information 15. Regulatory Information 16. Other Information

Following is a general description of the information contained in each section. 1. Identification. The trade name, product name, chemical formula and molecular weight where applicable, Chemical Abstracts Services (CAS) Registry Number, and chemical name many be found in this section. The chemical name is usually included only in the case of products being a single chemical. An emergency phone number is also given here for the purpose of obtaining further information if that is deemed necessary. 2. Ingredients information. Any ingredient contributing a hazard to the product must be given in this section along with its percentage or range of percentage and its CAS number. If the hazardous ingredient is a trade secret, a generic chemical name and a range of percentage may be used. In this case, the CAS number would be omitted since that is the same as giving the exact chemical me. Purity information is frequently given in this section for single chemical products. 3. Hazards identification. The primary hazards of the product, if any, are given here usually in the form of short terse warning and caution statements such as toxic, irritant, flammable. 4. First aid measures. First aid measures for various routes of exposure or contact are given here along with a caution to see a physician where appropriate. 5. Fire fighting measures. This section contains any special precautions to observe in the case of a fire. Flash point, formation of toxic gases, protective equipment, and information on appropriate extinguishing agents are given.

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6. Accidental release measures. This gives advice on how to handle a spill, such as special precautions for a material with low flash points, containment methods, and protective equipment. 7. Handling and storage. Some material should not be stored near incompatible materials. Highly flammable materials should be stored in well-ventilated areas such as outside. Some products may have flammable or toxic vapors in the head space. Information on these and similar concerns may be found here. 8. Exposure controls/personal protection. This section will advise on the need for special ventilation procedures, respirators, and other special equipment such as chemical splash goggles or special gloves or boots. 9. Physical and chemical properties. Those properties which aid in visual and other means of identification of the product and density in case the material gets into waterways are found here. 10. Stability and reactivity. Information pertaining to explosivity, thermal instability, and reactivity with other materials is placed in this section. 11. Toxicological information. This contains a summary of known toxicity information which may have been gleaned from the literature, resulted from laboratory testing on animals or tissue cultures, or observed with human exposure. 12. Ecological information. This section summarizes known effects on plant and animal life in the air, land, and water media. 13. Disposal considerations. This will advise what product characteristics should be considered before disposing of the product. In general, for liability reasons, highly specific guidance is not given. 14. Transport information. In general, facts such as the specific shipping classification, proper shipping name, and whether applicable to bulk or smaller shipments are listed. For the U.S., the road and rail designations and sometimes those for air shipments are given. Some companies include information for shipment by sea. 15. Regulatory information. There is wide variation regarding what is included here. One may expect to find the regulatory status under the U.S. Toxic Substances Control Act, and sometimes similar information for other countries. Components that may be affected by the U.S. Clean Air Act, Clean Water Act, Resource Conservation and Recovery Act, Superfund Amendments, and Reauthorization Act may also be noted here. 16. Other information. Other information which a supplier believes to be pertinent to health and safety but does not fit in the other sections may be given here.

TOXICOLOGY TESTING Most hazard warnings for lubricants and functional fluids are derived from in vitro tissue culture testing and testing in laboratory animals. Some, but very few, are derived from human experience. Tissue culture and animal testing are imperfect in that the correlations with human experience are not very good, but they are the best that we have to use. The testing of these fluids and the additive components that are used in these fluids present some unique problems. The neat materials (oil and solvent free) are too viscous in most cases for accurate dosing, so the bulk of the testing is conducted with material as manufactured, meaning usually with 30 to 60% oil present. Most of the substances have relatively high molecular weights, meaning that they may be too large to pass through animal or plant membranes. This would account for the low order of toxicity. Environmental hazard testing presents some unique problems, again based upon viscosity and molecular size. The materials having been designed to be soluble in oil have very low solubility in water and frequently do not mix well with water. For these reasons, much of the testing must be conducted with a water-accommodated fraction, that is, the portion of a nominal fraction that

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dissolves in water. Nominal concentrations are given in such cases due to difficulties in the analysis of these aqueous fractions.

ENVIRONMENTAL EXPOSURE Environmental exposures can occur by releases into the air; accidental spills onto soil or into sewers, streams, or bodies of water; or intentional dumping onto soil or into waterways. All of these are governed by one or more specific laws. Penalties for violation of specific provisions of these laws can be severe, particularly if the violation is willful. Willful violations can lead to prison terms. AIR RELEASES Release of lubricants or functional fluids or their individual components into the air is an uncommon occurrence since the components generally have low volatility, or they would not survive long enough to serve the intended functions. In some cases, low levels of volatile materials may be present and need attention due to the need to control workplace exposure as noted earlier. Releases that occur into the air surrounding a plant site are addressed, in the Clean Air Act. This Act established a list of hazardous air pollutants, substances which when released are anticipated to cause either mortality or serious illness. National Emission Standards for Hazardous Air Pollutants (NESHAPs) were established in Section 112(a) of the Act. The eight hazardous air pollutants listed in 40 CFR 61.01(a) are: arsenic, benzene, beryllium, coke oven emissions, inorganic arsenic, mercury, radio-nuclides, and vinyl chloride. The Environmental Protection Agency (EPA) is considering the addition of chromium, cadmium, and other hazardous organics and organic solvent cleaners to the list. It is worth noting that, in general, these listed materials are not used in the formulation of lubricants and functional fluids, but some may be present as trace impurities. WATER RELEASES Section 311(b)(2)(A) of the Clean Water Act (CWA) specifies reporting and response requirements for hazardous substances which, when released in an uncontrolled or unpermitted fashion, are deemed capable of causing environmental harm. Each substance listed has been assigned a reportable quantity (RQ) and requires immediate reporting when the release of a listed substance exceeds the RQ. This list may be found in 40 CFR 117.3. Another law, the Comprehensive Environmental Response, Compensation and Liability Act (CERCLA), expands the universe of hazardous substances and has its own reporting and response requirements. The list with RQ values may be found in 40 CFR 302.4. All substances from the CWA Section 311 list are found on the CERCLA list, but the reverse is not the case. SOIL RELEASES There is no law that specifically addresses releases which get into the soil, but provisions exist in other laws such as CERCLA discussed above. Significant releases which are not cleaned could result in the site being named a Superfund site.

SARA The Superfund Amendment and Reauthorization Act (SARA) significantly increased reporting requirements for all sites where chemicals or mixtures of chemicals may be used. This is essentially a right-to-know type of law with provisions for notification of local and state authorities of chemicals used at each site. Section 313 of this Act contains a list of more than 300 extremely hazardous substances by specific chemical name and CAS number. There is also a provision in Sections 311

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and 312 for notification of materials which have any of five hazardous properties, but which are not on the list of specific hazardous substances. These properties are: acute health hazards, chronic health hazards, sudden release of pressure, fire, and reactivity, which are condensed from the 23 OSHA categories of hazard. Under Sections 311, 312 owners or operators of certain facilities are required to provide local fire departments and local emergency planning commissions with quarterly updates and MSDSs or a list of MSDSs for all materials located at the site which meet the hazard requirements. Under Section 313, toxic chemical release forms must be used to file annual reports with U.S. EPA for the manufacture, processing, or use of any toxic material on the 313 list.4 In general, lubricants and functional fluids do not contain materials on the 313 list. Since the class of zinc compounds is listed, however, most automotive crankcase lubricants, some automatic transmissions fluids, and some hydraulic fluids may require attention due to their content of zinc alkyldithiophosphates. Certain lubricants may contain heavy metals which are listed. Any materials with irritation or skin sensitization properties or which may have low flash points, for example, may be subject to Sections 311 or 312.

WASTE DISPOSAL In general, fresh lubricants and functional fluids are not subject to disposal because of their cost. Buyers, for this reason, use them unless they become contaminated or have been spilled and are not recoverable. It is relatively easy to characterize the properties for waste disposal at this stage. After use, the task of characterization for waste disposal becomes more difficult due to changes and contamination that may have occurred during usage. In any event, disposal is governed by the Resource Conservation and Recovery Act (RCRA). Waste classified as hazardous must be treated, stored, transported, and disposed of in accordance with the applicable requirements of this Act. Many types of waste are addressed in this Act, but the intent here is only to address provisions which may apply to new and used lubricants and functional fluids. The primary goal of this Act is to ensure that all hazardous wastes are properly disposed of, either by incineration or placing in a secure land fill. The implementing regulations of the Act as given in 40 CFR 261 Subpart (C) assign four characteristics of hazardous waste: ignitable, corrosive, reactive, or toxic. Subpart (D) lists some wastes from nonspecific sources (Section 261.31) and specific sources (Section 261,32), as well as some discarded commercial chemical products, off-specification species, container residues, and spill residues (Section 261.33). Certain specific substances or constituents in a waste make it a hazardous waste and subject to all applicable provisions of RCRA. Although a fresh lubricant or functional fluid may not be classified as hazardous waste, changes which result from usage or contamination may make the used material a hazardous waste. It should be noted that used oil which contains in excess of 1000 ppm chlorine, 5 ppm arsenic, 2 ppm cadmium, 10 ppm chromium, 100 ppm lead or having a flash point of less than 100°F is considered to be hazardous waste and may not be burned for energy recovery or used in any way to produce fuel. Waste oil containing more than 1000 ppm chlorine is considered to be contaminated with chlorinated solvents or PCB and is designated as hazardous waste. This is subject to rebuttable presumption, meaning if it can be demonstrated that contamination with chlorinated solvents or PCB has not occurred, then the used oil is not hazardous waste (40 CFR Chapter I part 279). The rebuttable presumption provision does not apply to metalworking fluids containing chlorinated paraffins if they are processed through a tolling agreement to reclaim metalworking fluids. Classification as hazardous waste, of course, requires record keeping and special provisions for transportation, and severely limits possibilities for burning for fuel value. Aside from this provision, there are no constraints on burning used oil in any manner. Attention must also be given to specific state laws which are too numerous to discuss here. Some of these laws may contain additional constraints. Copyright © 1997 CRC Press, LLC.

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TRANSPORTATION The primary purpose of classifying chemicals and mixtures of chemicals for transportation is to provide information to all responsible parties about acute effects of the chemicals. The underlying assumption appears to be that truck drivers and railroad workers will have little or no contact with the material being transported, so the less severe acute health effects such as irritation and chronic health effects are not addressed in the classification scheme. The classifications focus on hazardous materials which are those having properties which can cause immediate and irreversible harm: poisons, flammables, corrosives, oxidizing materials, explosives, gases, radioactive substances, and agents capable of causing disease. These are the effects that are a concern in the event of a spill or an accident. The applicable law is the Hazardous Materials Transportation Act (HMTA). A list of 16,000 hazardous materials is found in 49 CFR 172.101. An appendix lists RQ values, relating to spills during transport, for substances which are also CERCLA hazardous substances. It should be noted that labeling according to the OSHA Hazard Communication Rule must be on containers when they leave work sites for transportation even though many of the hazards cited in the rule are not included in the HMTA. In shipping these regulated materials, care must be given for the proper preparation of shipping papers, labels, packaging, and vehicle (road or rail) placarding to give appropriate warning. In some cases different provisions apply to drum shipments and bulk shipments. The requirements for different modes of transportation may also differ and may be found for rail at 49 CFR 179, for aircraft (49 CFR 175), for vessels on waterways (49 CFR 176), and for road (49 CFR 177). Specifications for shipping containers may be found in 49 CFR 178 and for tank cars at 49 CFR 179. Storage of these regulated materials when received is addressed in the OSHA regulations at 29 CFR 1910 Subpart H.

OTHER APPLICABLE LAWS AND REGULATIONS FEDERAL LAWS The Toxic Substances Control Act (TSCA) governs directly or indirectly essentially all industrial chemicals manufactured, imported, or used in the U.S. While there are exemptions for a variety of materials covered by other laws, all components of lubricants and functional fluids are covered by this Act. The Act specifies that no chemical may be manufactured or used unless the chemical is on the TSCA Inventory of existing chemicals. Note that this is not a list of toxic chemicals but a list of existing chemicals, not subject to notification. A premanufacturing notification (PMN) (40 CFR 720) must be filed with EPA for any chemical not on this list at least 90 days before commencement of manufacture or importation. When manufacture or import of a chemical which is the subject of a PMN commences, the manufacturer or importer must notify EPA within 30 days. Special provisions apply to substances which will be manufactured in a quantity of less than 1 ton per year, certain polymers, and substances manufactured for export only (40 CFR 723). TSCA contains a number of provisions which can be a trap if not watched carefully. Use of a chemical which is not on the TSCA Inventory may be considered a violation, so it is prudent to check with a supplier to determine the compliance status. Section 4 of the Act (40 CFR 790) allows EPA to require testing on any chemical for which volume and presumed exposure is large and there are inadequate data to assess the risk potential of the substance. In general, these chemicals are made by more than one manufacturer, all of whom join together to conduct a joint testing program on the designated substance. In some cases, the testing provisions apply to processors. The Chemical Manufacturers Association under its CHEMSTAR programs provides a forum for such joint testing activities. Section 5 of TSCA (40 CFR 720) allows EPA to require testing on any new chemical for which a risk cannot be assessed. This testing delays the clearance of the PMN well beyond the usual 90

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days, sometimes to a year or two, and requires a considerable expenditure of money. Manufacture or import may not commence until such testing is complete and a report is in EPA’s hands. Section 6 (40 CFR 750) allows EPA to regulate manufacture, distribution, and use of any substance which in its judgment “presents an imminent and unreasonable risk of serious or widespread injury to health or the environment.” Such regulation may specify controls to limit exposure, limit the uses, or even ban the manufacture, import, and use of a substance. EPA currently regulates polychlorinated biphenyls (PCB), fully halogenated chlorofluoroalkanes, asbestos, and hexavalent chromium. Section 7 provides to EPA the authority to regulate chemicals which may pose an imminent hazard. It allows EPA to immediately stop sale or seize chemicals if deemed necessary. Chlorinated dibenzodioxins (Dioxin) and chlorinated dibenzofurans are regulated, but it is not clear if the action was taken under Section 6 or 7. These two classes of materials in general are of no concern in lubricants and functional fluids, but a concern could arise if these fluids contain chlorine and are burned. Section 8 contains several reporting provisions. Subsections (a) (40 CFR 704.1) and (d) require reporting certain data as specified by EPA in Federal Register notices. It is useful to review the Federal Register on a regular basis to ensure that such reporting requirements are not overlooked since the penalties for not reporting can be costly. Section 8(c) (40 CFR 717.1) requires record keeping for all allegations of potential health and environmental effects that are reported to a company which is using a chemical, providing the allegation can be attributed a specific chemical or product. Section 8(e) (43 Federal Register, page 11110, 1978) requires reporting within 15 working days information received which suggests that a chemical poses a substantial risk to human health or the environment. Information that is known to EPA, for example, published literature, need not be reported. A spill into a waterway which results in a large fish kill may, for example, be reportable under this provision in addition to any other reporting provision, as noted earlier. Ill feelings, irritation, skin sensitization may be reportable if not previously known. If in doubt, it is safer to report, since there is a penalty for not reporting and none for over-reporting. Section 12 (40 CFR 707.60) requires that the export of certain chemicals must be reported to EPA so that EPA may give the receiving country an option to deny the import of the chemical. Chemicals which are subject to Section 4 test rules, Section 5 consent orders or significant new use rules, chemicals listed in Section 6, and any chemicals that may be named under Section 7 are subject to this export notification. Export of certain chemicals deemed to be banned or severely restricted could become more onerous in the future, with growing international pressure for a legally binding instrument for prior informed consent and the potential for enlarging the list of chemicals to be covered. Section 13 (40 CFR 707.20) requires certification to the U.S. Customs authority that substances being imported comply with all applicable provisions of TSCA. STATE LAWS Because of the plethora of state laws and some local laws, no attempt will be made to be all-inclusive. This section will address some of the key issues. Many states have right-to-know laws which may require certain disclosures or the providing of MSDSs. It has been established that the OSHA Hazard Communication Rule preempts the state laws with regard to workplaces. Provisions that apply to community right-to-know still stand. California, in its Proposition 65, requires warnings on any product which contains an ingredient considered to be capable of causing cancer or birth defects. There is no concentration limit for this, making compliance difficult in many cases. The law has a bounty hunter provision. There is also a volatile organic compound reporting requirement under its SCAQMD Rule 443.1. Massachusetts requires the identification of any substance which is on its list of Hazardous Substances and Special Hazardous Substances which may be present in a product.

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New Jersey has a requirement for disclosing the five most prevalent components of a product in addition to any ingredients which contribute to the hazard of the product under the OSHA Rule. Trade secret protection for these ingredients is allowed. This applies only to companies conducting business in the state. Pennsylvania has a law very similar to that of Massachusetts.

BIODEGRADABILITY, DISPOSABILITY, RECYCLABILITY, AND LIFE CYCLE ANALYSIS These topics are related to the previous discussions on releases to the environment. Each of these topics fills volumes in the literature and the only purpose of addressing them here is to encourage awareness of the complicated nature of the ecosystem, better use of natural resources, and safe and responsible handling of all chemicals. BIODEGRADABILITY In order to reduce health and environmental concerns, more research is being focused on products which biodegrade in the environment. Products which may leak from machinery or marine vessels while being used are targeted for development of more biodegradable products. Fuels, hydraulic fluids, and two-cycle and four cycle engine oils that are biodegradable are now available. Biodegradability is often achieved by the use of naturally occurring seed oils. Although this results in an increase in the cost of the products, most petroleum-based oils do not meet current tests for biodegradability. The amount of research devoted to use of renewable resources in lubrication and other material areas has increased and may present significant challenges in the next decade. Some immediate issues involving lubrication include: new testing methods, standardization of current methods, development of more biodegradable lubricants, and new additive systems that are biodegradable. Perhaps the most urgent need today is in the area of test methodology and a clear definition of biodegradability with respect to lubricants. Biodegradability is generally judged on the basis of a ready biodegradability test (OECD Test Protocols 301A-F and 40 CFR 796.3100–796.3400) in water. Ready biodegradability means basically that the material decomposes to carbon dioxide and water to the extent of 60 to 70% within 28 days. The end points of these OECD methods vary, but 301-B measures carbon dioxide formation, which is what most regulatory authorities will accept today. It is recognized, however, that many substances degrade in the environment despite the fact that they are not degradable in water in this type of short-term testing. Decomposition in soil may differ from that in water. Also, bacteria which degrade materials sometimes need to acclimate to certain chemicals for a period of time before degrading them. Some lubricant users have begun to use the more recently developed CEC (Co-ordinating European Council) biodegradation test (“Biodegradability of Two-Stroke Cycle Outboard Engine Oils in Water,” CEC L-33-A-94, 1994). More lubricants decompose in this test than in the ready biodegradation tests. The test does not measure the formation of carbon dioxide, however, but measures the disappearance of a particular carbon-hydrogen bond in the infrared spectrum. More work needs to be done to gain a better understanding of what is important in the environment, so that appropriate tests can be used or developed to provide the best guidance. DISPOSABILITY The ability to properly dispose of waste materials, including used oil, whether they be hazardous chemicals, toxic, radioactive, or just plain garbage or solid wastes is one of the most critical problems related to the exploding growth and economic development of the world’s population. Municipal waste presents a problem because of diminishing areas suitable for landfilling. Industrial

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wastes from chemicals and chemical products present a problem because of concern over toxicity to the environment and a possible contamination of water supplies. Efforts aimed at finding alternatives, such as reduction in usage, recycling, burning, and improving biodegradability are expanding. A significant portion of used oil is burned for fuel value, some is recycled, and, unfortunately, some is released to the environment (soil, sewers). New ideas are critically needed to solve this problem. RECYCLABILITY Although major strides have been taken in a number of industries, recycling processes are sometimes difficult to develop, due to the complexity of the products. In addition, the cost of recycled materials may be higher, which limits the market. Some major users of metalworking fluids are believed to have efficient processes for breaking the emulsions of soluble oils to recover and recycle the oil. An option in addition to recycle is to reduce the amount of material for recycling or disposal by designing smaller units. This approach may increase the severity of the performance requirement and result in increased costs and longer product development times. Use of recycled products and biodegradable products is an area needing innovative thinking. A major deterrent to recycling of used oil in the U.S. is the absence of a comprehensive and convenient collection system. Some petroleum companies that operate gasoline stations and perform oil changing services and the fast-change lubricant stations collect oil and convert it for various uses. Some is known to go to oil-recycling plants and some is burned for its fuel value. A large portion of the used oil volume, however, is still in the do-it-yourself market. A major portion of this used oil may escape the recycling/reuse loop. Several years ago there was a proposal in a Senate committee for a law which would require the use of a certain percentage of recycled oil. That effort died and has not yet resurfaced. It is perhaps something that will surface again in the next few years. This undoubtedly would create a need for an efficient used-oil collection system and probably a need for additional re-refining capacity. LIFE CYCLE ANALYSIS The goal of life cycle analysis is to evaluate various options for production and use in terms of effects on health and the environment, raw material availability, waste generation, and disposal and cost. This will not lead to many early analyses, since it requires detailed knowledge and evaluation of raw materials, manufacturing processes, and how materials may be used and disposed of by the manufacturer and customer. Collecting and sorting through the many options is a long, tedious, and expensive process which could require many months or even years for completion of the analysis for a single chemical. Another major problem in life cycle analysis is lack of consensus on how to assess the relative value of various options. For example, is it better to reduce waste which goes to a landfill, to burn waste, or to use more energy or different raw materials in the manufacturing process? These are all options for the manufacture of the same material. Burning of waste or use of more energy in the manufacturing process may be considered by some as contributors to global warming, whereas the first option and last options could potentially lead to ground water contamination, human contact in future generations, or other environmental effects. There are no easy solutions to these difficult problems.

CONCLUSIONS The increased awareness of health, safety, and environmental issues surrounding lubricants and functional fluids has led to products having significantly lower health and environmental risks for the users. Hopefully, awareness of these issues by the end users has also increased. This progress

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will continue, while at the same time producing products with improved properties for the intended uses. Because of these concerns the manufacturers and users of these products must act responsibly and must pay close attention to the laws and regulations which have been developed over the years to address these concerns.

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REFERENCES 1. Ignatowski, A.J., Hamilton, J.D., and Weiler, E.D., Regulatory Toxicology and Pharmacology, 22, 231–242, 1995, Review of international criteria and mixture rules for health hazard classification, published by Academic Press. 2. Cisson, CM., Raussina, G.A., and Stonebraker, P.M., Human health and environmental hazard characterization of lubricating oil additives, Tribology Coloquium, Esslingen, Germany 1994.

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3. Rausina, G., Barth, M., Blackmon, J., Hoke, D., Isola, D., Linnett, S., Ribero, P., Stack, C, and Ward, T., Aquatic toxicity of petroleum additives, SETAC Conf., November 1995 and Tribology Conference, Esslinger, Germany, January 1996, submitted. 4. Toxic Chemicals Subject to Section 313 of the Emergency Planning and Community Right-toKnow Act of 1986, Committee Print No. 99–169, Senate Committee on Environment and Public Works.

INFORMATION SOURCES INTERNET Federal Register (1995 and later only) http://www.access.gov.gov/su_docs/aces/aces140.html International Agency for Research on Cancer Homepage: http://www.iarc.fr U.S. Environmental Protection Agency (full text of regulations not yet available) Homepage: http://www.epa.gov 40CFR contents: http://www.epa.gov/epacfr/40 Occupational Safety and Health Administration Homepage: http://www.osha.gov 29CFR 1910.1200: http://oshstd_data/1910.1200.html (full text available) AGENCIES AND PUBLICATIONS American Conference of Governmental Industrial Hygienists Technical Affairs Office 1330 Kemper Meadow Drive Cincinnati, OH 45240 American National Standards Institute 11 W. 42nd. St. New York, NY 10036 Code of Federal Regulations — individual sections available from the Superintendent of Documents. Department of Transportation Research and Special Programs 400 Seventh St., S.W. Washington, D.C. 20590 (phone: 202-366-4000) Environmental Protection Agency Office of Toxic Substances 401 M St. S.W. Washington, D.C. Federal Register Superintendent of Documents Government Printing Office Washington, D.C. 20402 (phone: 202-783-3238) (May also be available in some public libraries, see also Internet above.)

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International Agency for Research on Cancer (publications) World Health Organization Distribution and Sales Service 1211 Geneva 27 Switzerland

WHO Publications Center 49 Sheridan Ave. Albany, NY 12210

Occupational Health and Safety Agency 200 Constitution Avenue Washington, D.C. 20210 (phone: 202-523-7162) OECD Publications and Information Center 2001 L St., N.W., Suite 700 Washington, D.C. 20036-4910 Registry of Toxic Effects of Chemical Substances (several volume set), S/N 17-33-00431-5. Available from the Superintendent of Documents. State Environmental Protection Agencies State capital cities Trade Associations American Petroleum Institute 1220 L St. Northwest Washington, D.C. 20005-4070 Chemical Manufacturers Association 1300 Wilson Blvd. Arlington, VA 22209 Independent Lubricant Manufacturers Association 651 So. Washington St. Alexandria, VA 22314 National Association of Oil Recovery Coordinators 325 North Adams Springfield, IL 62706 U.S. Code: available from Superintendent of Documents.

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Appendixes

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Appendix 1: Material Hardness Tables, Tests and Data Charles A. Moyer Hardness is a property of a material based, in general, on its resistance to plastic flow. Hardness is determined using tests that try to measure indentation, abrasion (or scratching), ploughing, cutting, rebound (bounce height), dampening (pendulum amplitude), or erosion (grain or droplet impingement). Which test is used depends on the material and application needs.1 The tables presented in this section are based primarily on indentation tests that account for most hardness testing methods. However, rebound devices are used for measuring hardness on surfaces that cannot tolerate surface marking or are limited by size. The theory behind indentation hardness testing is covered in Reference 1 and additional background is included in References 2 and 3. The tables list Vickers,4 Brinell,5 Rockwell,6 Knoop,7 Shore,8 and approximate tensile strength to match the various hardness numbers. Although tensile strength is given in psi, the values given in the columns can be converted to MPa by multiplying by 6.895 (i.e., Table 2, 355 × 6.895 = 2450 MPa). Assuming a relation between load applied and indentation, then hardness is measured by indent surface area or depth. Table 1 (from Reference 1) is a summary of indenters, indent size, loads measurement form, surface preparation, and applications. Tables 2, 3, and 4 are from Reference 2. These provide comparison values for various indentation hardness tests that can be made. Table 5 is summarized from Reference 9, so the range of hardness for the Moh scale can be included. This table also has approximate values for various elements and other materials, and some specific values from the Knoop scale to provide some relationships to the other tables.

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REFERENCES 1. Hardness testing, in ASM Metals Handbook, 9th ed., vol. 8, Mechanical Testing, ASM International, Metals Park, OH, 1985, 69–113. 2. Hardness testing, in ASM Metals Handbook, 8th ed., vol. 11, Nondestructive Inspection and Quality Control, ASM International, Metals Park, OH, 1976, 1–20. 3. Shaw, M. C. and DeSalvo, G. J., On the plastic flow beneath a blunt axisymmetric indenter, Trans. ASME, J. Eng. Industry, May, 1970, 480–494. 4. Standard Test Method for Vickers Hardness of Metallic Materials, E 92, Annual Book of ASTM Standards, American Society of Testing and Materials, Philadelphia, 1984, 253–263. 5. Standard Test Method for Brinell Hardness of Metallic Materials, E 10, Annual Book of ASTM Standards, American Society of Testing and Materials, Philadelphia, 1984, 164–169. 6. Standard Test Methods for Rockwell Hardness and Rockwell Superficial Hardness of Metallic Mate rials, E 18, Annual Book of ASTM Standards, ASTM, Philadelphia, 1984. 7. Standard Test Methods for Microhardness of Materials, E 384, Annual Book of ASTM Standards, ASTM, Philadelphia, 1984, 497–518. 8. Miscellaneous hardness testing, ASM Metals Handbook, 9th ed., vol. 8, Mechanical Testing, ASM International, Metals Park, OH, 1985, 104–106. 9. Lide, D. R., Ed., Handbook of Chemistry and Physics, 75th ed., CRC Press, Boca Raton, FL, 1994, p. 12–186.

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Appendix 2 Viscosity Conversion Factors Compiled by Douglas Godfrey Note: “E” (exponent) implies 10 raised to a power, for example: 2.000E + 03 = 2.000 × 10+3 = 2000 2.000E - 03 = 2.000 × 10-3 = 0.002 (f) = force.

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BIBLIOGRAPHY ON STYLE Standard Practice for Use of the SI International System of Units, The Modern Metric System, ASTM E380–93, American Society for Testing and Materials, Philadelphia, 1993, 30–31. ON VISCOSITY CONVERSIONS 1. Viscosity conversion, in CRC Handbook of Chemistry and Physics, 62nd ed., Weast, R. C. and Astle, M. J., Eds., 1981–1982, CRC Press, Boca Raton, FL, 1981-1982, F-41 [extensive list]. 2. CRC Handbook of Chemistry and Physics, 72nd ed., CRC Press, Boca Raton, FL, 1994. 3. Standard Practice for Conversion of Kinematic Viscosity to Saybolt Universal Viscosity or to Saybolt Furol Viscosity, ASTM Designation D 2161–93, in ASTM 05.01 Petroleum Products and Lubricants D56-D2596, ASTM, Philadelphia. 4. Alexander, D. L., Viscosity of lubricants, L ubrication, Vol. 78, No. 3, Texaco Inc., New York, 1992. 5. “Viscosity,” L ubrication, Vol. 52, No. 4, Texaco, Inc., New York, 1966. 6. Lindeburg, M. R., E ngineering Unit Conversions, 3rd ed., Professional Publications, Belmont, CA, 1993 [156 entries of viscosity conversions].

OTHER VISCOSITY UNIT CONVERSIONS* A viscosity determined in a particular viscometer at a specific temperature can be converted to the equivalent viscosity which would have been obtained had the liquid been tested in some other viscometer at the same temperature. For the most accurate conversion, the determination temperature must be considered. Some of the wide variety of equations, tables, charts, and nomographs which have been developed to facilitate such conversions are described in the following. KINEMATIC VISCOSITY TO SAYBOLT UNIVERSAL AND FUROL ASTM has recently issued a method, D 2161–63T, giving tables in a convenient form and equations for converting kinematic viscosity in centistokes at any temperature to Saybolt Universal seconds (SUS) at the same temperature. Also, tables are included for converting kinematic viscosity at 122 or 210°F to Saybolt Furol seconds at the same temperatures. A supplement to method D 2161 was recently issued. This consists of a set of tables giving the viscosity of a petroleum oil in Saybolt Universal seconds at 0°F obtained by extrapolating measured kinematic viscosities at 100 and 210°F down to 0°F. Extrapolated values are based on the assumption that a strictly linear relationship exists between viscosity and temperature on the ASTM Viscosity–Temperature charts of ASTM Method D 341. Kinoshita** developed the following equation for converting kinematic viscosity v in centistokes, determined at temperature t (°F) into Saybolt Universal seconds (SUS).

This equation is satisfactory for kinematic viscosities greater than 2 cSt and temperatures between 70 and 300°F.

* Courtesy of Texaco magazine Lubrication, Vol. 52, No. 3, Texaco, Inc., New York, 1966. With permission. ** Kinoshita, M., J. Inst. Pet., 43, May 1957.

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KINEMATIC TO REDWOOD NO. 1 (STANDARD) The following Kinoshita* equation for this conversion is highly accurate for kinematic viscosities above 3 centistokes and temperatures between 70 and 210°F.

KINEMATIC VISCOSITY TO ENGLER DEGREES A table in the German Standard DIN 51 560, “Determination of Viscosity with the Engler Apparatus,” gives the relationship between kinematic and Engler (degrees) viscosities. CONVERSION OF “SHORT-TUBE” VISCOSITIES TO KINEMATIC VISCOSITY The following five groups of equations provide approximate conversions of “short-tube” viscosities to kinematic viscosities at the same temperature. In all cases “T” is the efflux time in seconds and “v” is the desired kinematic viscosity in centistokes. SAYBOLT UNIVERSAL SECONDS TO KINEMATIC

SAYBOLT FUROL SECONDS (SFS) TO KINEMATIC

REDWOOD NO. 1 (STANDARD) SECONDS TO KINEMATIC

REDWOOD NO. 2 (ADMIRALTY) SECONDS TO KINEMATIC

ENGLER DEGREES (E°) TO KINEMATIC

* Kinoshita, M., J. Inst. Pet., 43, May 1957.

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FIGURE 1 Viscosity conversion nomograph. Line up straight edge so centistoke value on both kinematic scales is the same. Viscosities at the same temperature on all scales are then equivalent. To extend range of only the kinematic, Saybolt Universal, Redwood No. 1 and Engler scales: multiply by 10 the viscosities on these scales between 100 and 1000 centistokes on the kinematic scale and the corresponding viscosities on the other 3 scales. For further extension, multiply these scales as above by 100 or a higher power of 10. (Example: 1500 centistokes = 150 × 10 cSt ≅ 695 × 10 SUS = 6950 SUS.) (Courtesy of Texaco magazine Lubrication, Vol. 52, No. 4, 1966.)

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FIGURE 2 Relationship between kinematic viscosity (cSt) and Saybolt Universal Seconds (SUS). (Courtesy of Texaco magazine Lubrication, Vol. 78, No. 3, 1992.)

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Appendix 3 International System of Units (SI) and Conversion Factors*

The International System of units (SI) was adopted by the 11th General Conference on Weights and Measures (CGPM) in 1960. It is a coherent system of units built from seven SI base units, one for each of the seven dimensionally independent base quantities: they are the meter, kilogram, second, ampere, kelvin, mole, and candela, for the dimensions length, mass, time, electric current, thermodynamic temperature, amount of substance, and luminous intensity, respectively. The definitions of the SI base units are given below. The SI derived units are expressed as products of powers of the base units, analogous to the corresponding relations between physical quantities but with numerical factors equal to unity. In the International System there is only one SI unit for each physical quantity. This is either the appropriate SI base unit itself or the appropriate SI derived unit. However, any of the approved decimal prefixes, called SI prefixes, may be used to construct decimal multiples or submultiples of SI units. It is recommended that only SI units be used in science and technology (with SI prefixes where appropriate). Where there are special reasons for making an exception to this rule, it is recommended always to define the units used in terms of SI units. This section was reprinted with the permission of IUPAC.

DEFINITIONS OF SI BASE UNITS Meter — the meter is the length of path travelled by light in vacuum during a time interval of 1/299 792 458 of a second (17th CGPM, 1983). Kilogram — The kilogram is the unit of mass; it is equal to the mass of the international prototype of the kilogram (3rd CGPM, 1901). Second — The second is the duration of 9 192 631 770 periods of the radiation corresponding to the transition between the two hyperfine levels of the ground state of the cesium-133 atom (13th CGPM, 1967). Ampere — The ampere is that constant current which, if maintained in two straight parallel conductors of infinite length, of negligible circular cross-section, and placed 1 meter apart in a vacuum, would produce between these conductors a force equal to 2 × 10-7 newton per meter of length (9th CGPM, 1948).

* These descriptions and conversion factors are taken from the CRC Handbook of Chemistry and Physics, Lide, D. R., Ed., 76th ed., (1995– 1996), pp. 1– 20 to 1– 39, 1995.

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Kelvin — The kelvin, unit of thermodynamic temperature, is the fraction 1/273.16 of the thermodynamic temperature of the triple point of water (13th CGPM, 1967). Mole — The mole is the amount of substance of a system which contains as many elementary entities as there are atoms in 0.012 kilogram of carbon-12. When the mole is used, the elementary entities must be specified and may be atoms, molecules, ions, electrons, other particles, or specified groups of such particles (14th CGPM, 1971). Examples of the use of the mole: 1 mol of H2 contains about 6.022 × 1023 H2 molecules, or 12.044 × 1023 H atoms 1 mol of HgCl has a mass of 236.04 g 1 mol of Hg2Cl2 has a mass of 472.08 g 1 mol of Hg2has a mass of 4.01.18 g and a charge of 192.97 kC 1 mol of Fe091 S has a mass of 82.88 g 1 mol of e- has a mass of 548.60 µg and a charge of -96.49 kC 1 mol of photons whose frequency is 1014 Hz has energy of about 39.90 kJ Candela — The candela is the luminous intensity, in a given direction, of a source that emits monochromatic radiation of frequency 540 × 1012 hertz and that has a radiant intensity in that direction of (1/683) watt per steradian (16th CGPM, 1979). NAMES AND SYMBOLS FOR THE SI BASE UNITS

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SI PREFIXES To signify decimal multiples and submultiples of SI units the following prefixes may be used.

Prefix symbols should be printed in roman (upright) type with no space between the prefix and the unit symbol. Examples

kilometer, km

When a prefix is used with a unit symbol, the combination is taken as a new symbol that can be raised to any power without the use of parentheses. Examples

1 cm3 = (0.01 m)3 = 10-6 m3 1 µs-1 = (10-6 s)-1 = 106 s-1 1 V/cm = 100 V/m 1 mmol/dm3 = mol m-3

A prefix should never be used on its own, and prefixes are not to be combined into compound prefixes. Example

ppm, not mmm

The names and symbols of decimal multiples and sub-multiples of the SI base unit of mass, the kg, which already contains a prefix, are constructed by adding the appropriate prefix to the word gram and symbol g. Examples

mg, not mkg; Mg, not kkg

The SI prefixes are not to be used with °C.

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UNITS IN USE TOGETHER WITH THE SI These units are not part of the SI, but it is recognized that they will continue to be used in appropriate contexts. SI prefixes may be attached to some of these units, such as milliliter, ml; millibar, mbar; megaelectronvolt, MeV; kilotonne, ktonne.

REFERENCES Quantities, Units and Symbols in Physical Chemistry, Ian Mills, Editor, Blackwell Scientific Publications, Oxford, 1987. Symbols, Units, Nomenclature and Fundamental Constants in Physics, Cohen, E. R. and Giacomo, P., Eds., International Union of Pure and Applied Physics Document 25, 1987; also appears in Physica 146A, 1–8, 1987. ISO Standards Handbook 2, Units of Measurement, International Organization for Standardization, Geneva, 1982. OTHER REFERENCES: Standard Practice for the Use of the International System of Units (SI) — (The Modernized Metric System), ASTM E 380-93, American Society for Testing and Materials, Philadelphia, PA. [Includes 8 pages of units classified by physical quantity, such as pressure.] Lindeburg, M. R., Engineering Unit Conversions, 3rd ed., Professional Publications, Belmont, CA, 1993 [148 pages of unit conversions].

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CONVERSION FACTORS The following table gives conversion factors from various units of measure to SI units. It is reproduced from NIST Special Publication 811, Guide for the Use of the International System of Units (Superintendent of Documents, U.S. Government Printing Office, 1991), which in turn was derived from IEEE Std 268 — 1982, IEEE Standard Metric Practice (© 1982 by the Institute of Electrical and Electronics Engineers, Inc.). The SI values are expressed in terms of the base, supplementary, and derived units of SI in order to provide a coherent presentation of the conversion factors and facilitate computations (see the table “International System of Units” in this Section). Powers of ten can be avoided by using SI prefixes and shifting the decimal point if necessary. Conversion from a non-SI unit to a different non-SI unit may be carried out by using this table in two stages, e.g., 1 cal (thermochemical) = 4.184 J 1 Btu (mean) = 1.05587 E+03 J Thus, 1 Btu (mean) = (1.05587 E+03/4.184) cal (thermochemical) = 252.359 cal (thermochemical) Conversion factors are presented for ready adaptation to computer readout and electronic data transmission. The factors are written as a number equal to or greater than one and less than ten with six or less decimal places. This number is followed by the letter E (for exponent), plus or minus symbol, and two digits which indicate the power of 10 by which the number must be multiplied to obtain the correct value. Examples: 3.523 907 E-02 is 3.523 907 × 10-2 3.386 389 E+03 is 3.386 389 × 103 An asterisk (*) after the sixth decimal place indicates that the conversion factor is exact and that all subsequent digits are zero. All other conversion factors have been rounded to the figures given in accordance with accepted practice. Where less than six decimal places are shown, more precision is not warranted.

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