Handbook of Lubrication Episode 2 Part 8 pptx

field testing and field development of lubrication requirements for a particular equipment installation are often necessary. Bearings and bearing units are designed for service ranging from nonregreasable (lubri- cated-for-life) to almost continuous relubrication by means of automatic systems. Advantages of grease over oil lubrication include the ease of sealing it within the bearing, the ability of grease to seal out contaminants, and its ability to coat parts and provide good corrosion protection. Disadvantages of grease include its inability to remove heat or flush away wear products, the possibility of accumulating dirt or other abrasive contamination, and a potential incompatibility problem if thickeners of different types are mixed. Oils Oil can be pumped, circulated, filtered, cleaned, heated, cooled, and atomized. Its ad- vantages over grease include its ability to remove heal, flush away wear products and contaminants, and to be recycled. It is more versatile than grease and is suitable for many severe applications involving extreme speeds and high temperatures. On the other hand, it is more difficult to seal or retain in bearings and housings. Oil level or oil flow in high- speed bearings is critical and must be properly controlled. Selection of proper oil viscosity is essential and is based primarily on expected operating temperature, speed, and bearing geometry. Excessive oil viscosity many cause skidding of rolling elements and undue lubricant friction with severe overheating and raceway damage. Insufficient oil viscosity may result in metal contact and possible premature failure. Other oil properties such as viscosity index, flash point, pour point, neutralization number, carbon residue, and corrosion protection are of varying significance in specific installations. Synthetic Lubricants Development of synthetic lubricants was initially prompted largely by the extreme en- vironmental demands of military and aerospace activities. Currently the following classes of synthetic oils are available as bearing lubricants: (1) synthetic hydrocarbons, such as alkylated aromatics and olefin oligomers, (2) organic esters, such as dibasic acid esters, polyol esters and polyesters, (3) others, such as halogenated hydrocarbons, phosphate esters, polyglycol ethers, polyphenyl ethers, silicate esters, and silicones, and (4) blends, which would include mixtures of any of the above. Use of a synthetic lubricant in a commercial application may be dictated by extreme operating conditions, for fire resistance, to meet a specification or code requirement, or to conserve petroleum-based lubricants. Although synthetic lubricants usually permit a much broader operating temperature range, temperature limits for synthetics are often misunder- stood. For example, in various aircraft and space applications operation at extremely high temperature is essential but life requirements may be very short. Since industrial requirements are usually for much longer periods of operation, temperature limits for a given synthetic in industry can be much lower. Some synthetic lubricants may also have other limiting characteristics such as in load-carrying ability and high-speed operation. Dry Lubricants Dry, or solid lubricants are usually used under conditions of high temperature or where boundary lubrication prevails. For example, notable success has been achieved by solid lubrication of kiln car wheels, conveyor wheels, and furnace roll bearings. These high- temperature applications involve extremely low speed where ample torque is available to rotate the bearing at a relatively high coefficient of friction. Solid lubricants may simply be dusted as a dry powder on parts to be lubricated, or they may be placed in a liquid carrier. The liquid may either be a fluid intended to evaporate or it may itself be a lubricating liquid or grease. Solid films are also applied as a bonded 528 CRC Handbook of Lubrication Copyright © 1983 CRC Press LLC coating. Some of the more common materials used are graphites, molybdenum disulfide, cadmium iodide, and fluorinated polyethylenes. Typical bonding agents are resins, silicone, ceramics, and sodium silicate. Another method incorporates the lubricant into one or more of the bearing components, typically a bearing retainer. Soft metals such as silver and tin could be used for this process. In such cases the dry lubricant is transferred from the cage to the bearing raceways by the rolling elements rubbing against the cage. Bearing life is governed by the wear-out life or depletion of the lubricant. Since these special bearings are usually quite expensive, practical industrial practice is to design equipment for use of conventionally lubricated bearings. Lubricant Temperature Limits Temperature is the major factor affecting life of a rolling bearing lubricant. Lubricant temperature is influenced primarily by bearing speed, bearing load, ambient temperature, and lubricant system design. With two different greases used on identical applications, base oil type and viscosity, thickeners, and chemical structure can all contribute to different operating temperatures. Some greases will churn in high-speed bearings and cause over- heating, whereas a channeling type grease may function satisfactorily at a much reduced temperature. Extremely low temperatures must also be considered. The lubricant must permit an ac- ceptable starting torque and must not freeze or become too stiff. While the lubricant must permit equipment turnover at the lowest temperature, it must also have adequate viscosity at the higher operating temperatures to provide sufficient oil film strength. For example, a petroleum type lubricant with very low viscosity oil considered for startup at –40°C and operation at 40°C may be unsuitable for operation at 80°C. In such cases, a synthetic oil or grease may be required to function satisfactorily at both the high and low limits. Tables 15 and 16 give approximate operating temperature limits for greases and oils. As mentioned previously, however, performance can vary widely depending upon the specific details of a given application. Additives can also affect the suitable operating temperature limits. They can, for example, be somewhat extended by oxidation-inhibiting additives or they may be somewhat reduced by EPor antiwear additives. Earlier chapters of this hand- book, along with References 27 through 29, provide more detailed information on various lubricant factors. Consultation with a reputable lubricant supplier is highly recommended. Lubricant Selection Table 17 illustrates “critical” ranges of extreme load, speed, or temperature where special Volume II529 Table 15 TYPICAL OPERATING TEMPERATURE LIMITS FOR GREASES Copyright © 1983 CRC Press LLC brication. Visual gages are usually provided to facilitate checking for a continuous lubricant supply to all bearings in the system. In cases where separate bearings operate under different conditions of temperature, speed, and load, use of more than one system may be necessary to meet the correct lubrication needs of the individual bearings. Circulating oil lubrication systems are most beneficial when bearings must be cooled continuously and when abrasive materials must be flushed away to assure safe operation. Circulating oil lubrication systems nearly always have filter and heat exchanger elements in addition to their oil reservoir and pump. They may also have a centrifuge or a sump for separating and removing foreign material, remote controls, warning devices, automatic cut- off switches, etc. These are particularly useful in meeting the special requirements of paper mills, lumber mills, steel mills, coal processing plants, and similar applications. Oil mist lubrication systems use an air stream to provide oil to the bearings. The air pressure maintains a positive pressure within the bearing chamber which effectively prevents foreign matter from entering. The air flow can be regulated to produce minimum lubricant friction and the concomitant lubrication friction temperature effect. The air flow will not, however, provide significant cooling. Air flowing out of a mist-lubricated bearing may discharge a fine oil vapor. This vapor may be objectionable, especially in the food and textile industries. In such cases, it is necessary to vent to other areas or provide air cleaning systems. Drainage of bearing res- ervoirs, provision for proper oil levels during bearing start-up, and timing of the mist flow must meet precise specifications. For this reason the system manufacturer should be relied upon to adjust the system for correct operation. Detailed information on lubricating systems is given in other chapters. FAILURE ANALYSIS Selection, application, and installation of rolling element bearings is based on subsurface nucleated fatigue. In the field, however, only 5 to 10% of the bearings removed from service are found to have developed this type of failure such as illustrated in Figure 23. 532 CRC Handbook of Lubrication FIGURE 23. Subsurface nucleated spall on cylindrical bearing inner ring raceway. (Magnification × 50.) Copyright © 1983 CRC Press LLC Volume II 533 Table 18 FAILURE MODES THAT LIMIT PERFORMANCE Copyright © 1983 CRC Press LLC Fatigue can often be induced by maldistribution of load in bearings due to varying stiffness of the mounting or support surfaces, housings, or shafts. Recognizing the sensitivity of rolling element bearing life to the variations in stress under the most heavily loaded rolling element (ball bearing life ~ (1/Stress) 8-10 , roller bearing life ~ (1/Stress) 7-9 ), the designer must carefully consider the mounting, its stiffness, and the influence of mutual deflections of all components in the system. Distortions due to temperature distributions are equally important and transient conditions must be properly accounted for. Damage commonly results from imposed loads which differ considerably from those anticipated in a machine design. Misalignment or fitting errors in mounting a bearing, misalignment or coupling faults between two machines, differential thermal expansion in a frame and shaft system, and rotor unbalance are among such factors. Simple visual or low power microscopic analysis of the ball paths in a ball bearings will frequently enable a useful evaluation of the magnitude and nature of these operating conditions. 30 Table 18 lists failure modes that limit the performance of rolling element bearings. Several bearing companies have published similar lists and several volumes have been written on the subject. Of particular note is Reference 31. Detailed failure analysis should be correlated with the bearing company involved since their laboratory, background, and experience enable them to draw conclusions and make recommendations. 534 CRC Handbook of Lubrication Table 18 (continued) FAILURE MODES THAT LIMIT PERFORMANCE FIGURE 24. Scanning electron micrograph of surface nucleated spall. Copyright © 1983 CRC Press LLC Bearings which have been grease-lubricated require special attention since they generally show surface effects which are the combination of many operating regimes. Grease lubri- cation in many bearings is a variable which depends on frequency of relubrication or on cyclic temperature variations to which the bearing is exposed. Many greased bearings operate with depleted films and significant wear will obliterate many original evidences of loading. Caution must be exercised against relying on the obvious conclusions while paying insuf- ficient attention to the minor findings evidenced on careful examination. Failure analysis has been greatly aided by utilization of the scanning electron micro- scope. 31,32 Small differences in the surface can indicate either the immediate condition of the bearing or some condition which has resulted from its previous operation. Figures 24 to 27 show scanning electron micrographs of bearing components which had been removed from service. Equally important is analysis of lubricants which can indicate the extent of contamination and deterioration. More recently, “ferrographic” analysis of filtrants or tail- ings from lubricants has become a valuable tool in monitoring transient bearing condition. Since catastrophic failures have reduced value in aiding the troubleshooter, every effort must be made to look at units which have not failed completely, preferably a number of them with different periods of operation in order to detect and trace the incipient failure mode. Of particular importance is the observation of changes in surfaces, the lubricant, housing, and shaft as well as the bearing to correct outside influences that can cause early bearing failure. In many instances, misalignment of sufficient magnitude to cause moment loading to run the rolling elements off the raceway induces a violent premature fatigue failure. Obviously, severe misalignment of tapered and cylindrical bearings must be corrected initially. Practical limits for misalignment are shown in Table 13. Volume II535 FIGURE 25.SEM of ground surface after running. (Magnifications × 100 and 500.) Copyright © 1983 CRC Press LLC REFERENCES 1. Metric Ball and Roller Bearings Conforming to Basic Boundary Plans, ANSI/AFBMA Standard 20, Anti- Friction Bearing Manufacturers Association, Arlington, Va., 1977. 2. Tapered Roller Bearings — Radial Inch Design, ANSI/AFBMA Standard 19, Anti-Friction Bearing Man- ufacturers Association, Arlington, Va., 1974. 3. Boresi, A. P., Sidebottom, O. M., Seely, F. B., and Smith, J, O., Advanced Mechanics of Materials, 3rd ed, John Wiley & Sons, New York, 1978, 581. 4. Harris, T. A., Rolling Bearing Analysis, John Wiley & Sons, New York, 1966. 5. Palmgren, A., Ball and Roller Bearing Engineering, SKF Industries, Philadelphia, 1959. 6. Eshmann, Hasbargen, and Weigand, Ball and Roller Bearings, Their Theory, Design, and Application, K. G. Heyden & Co., London, 1958. 7. Hartnett, M. J., The analysis of contact stresses in rolling element bearings, ASME Trans. J. Lubr. Technol. 101(1), 105, 1979. 8. Hamrock, B. J., Stresses and Deformations in Elliptical Contacts, Tech. Memo. 81535, National Aero- nautics and Space Administration, Washington, D.C., 1981. 9. Lundberg, G., Cylinder Compressed Between Two Plane Bodies, Aktiebolaget, Svenska Kullagerfabriken, Goteborg, 1949. 10. Load Ratings and Fatigue Life for Ball Bearings, ANSI/AFBMA Standard, Anti-Friction Bearing Manu- facturers Association, Arligton, Va., 1978. 11. Load Ratings and Fatigue Life of Roller Bearings, ANSI/AFBMA Standard II, Anti-Friction Bearing Manufacturers Association, Arlington, Va., 1978. 12. Lundberg, G. and Palmgren, A., Dynamic capacity of rolling bearings, Acta Polytech. Mech. Eng. Ser., 1 (3), 1952. 13. Lundberg, G. and Palmgren, A., Dynamic capacity of roller bearings, Acta Polytech. Mech. Eng. Ser., 2(4), 1952. 14. Moyer, C. A. and McKelvey, R. E., A rating formula for tapered roller bearings, SAE Trans., 71, 490, 1963. 15. Price, C. E. and Galambus, M., Bearing Application for Material Conveying Equipment, Paper No. 80- 3011, American Society of Agricultural Engineers, St. Joseph, Mich., 1980. 16. Grubin, A. N. and Vinogradova, I. E., Investigation of the contact of machine components, TsNIITMASh, book No. 30, Department of Scientific and Industrial Research, London, 1949. 17. Dowson, D. and Higginson, G. R., A numberical solution to the elastohydrodynamic problem, J. Mech Eng.Sci., 1(1), 6, 1959. 18. Puckett, S. J. and Pflaffenberger, E. E., Rolling Contact Bearing Surfaces — The Current Relationship Between Requirements and Processing, Paper No. 1073, Society of Manufacturing Engineers, Dearborn, Mich., 1973. 19. Tallin, T. E., On competing failure modes in rolling contact, ASLE Trans., 11, 418, 1967. 20. Littmann, W. E., Widner, R. L., Wolfe, J. O., and Stover, J. D., The role of lubrication in propagation of contact fatigue cracks, ASME Trans. J. Lubr. Technol. Ser. F, 90(1), 89, 1968. 21. Rounds, F. G., Some effects of additives on rolling contact fatigue, ASLE Trans., 10, 243, 1967. 22. Bock, F. C., Bhattacharyya, S., and Howes, M. A. H., Equations relating contact fatigue life to some material, lubricant, and operating variables, ASLE Trans., 22(1), 1, 1979. 23. Life Adjustment Factors for Ball and Roller Bearings, An Engineering Design Guide, American Society of Mechanical Engineers, New York, 1971. 24. Danner, C. H. Fatigue life of tapered roller bearings under minimal lubricant films, ASLE Trans., 13, 241, 1970. 25. Hamrock, B. J. and Dowson, D., Isothermal elastohydrodynamic lubrication of point contacts. III. Fully flooded results, ASME Trans. J. Lubr. Techno., 99(2), 264, 1977. 26. Coy, J. J. and Zaretsky, E. V., Some limitations in applying classical EHD film thickness formulae to a high speed bearing, ASME J. Lubr. Technol., 103(2), 295, 1981. 27. Neale, M. J., Ed., Tribology Handbook, John Wiley & Sons, New York, 1973. 28. Szeri, A. Z., Ed., Tribology — Friction, Lubrication, and Wear, Hemisphere Publishing, Washington, D.C., 1980. 29. Hatton, R. E., Synthetic oils, in Interdisciplinary Approach to Liquid Lubricant Technology, NASA SP- 318, National Aeronautics and Space Administration, Washington, D.C., 1973. 30. ASLE Manual, Interpreting Service Damage in Rolling Type Bearings, American Society of Lubrication Engineers, Park Ridge, III., 1953. 31. Tallian, T. E., Baile, G. H., Dalal, H., and Gustafson, O. G., Rolling Bearing Damage Atlas, SKF Industries, King of Prussia, Pa., 1974. 32. Derner, W. J., The Use of the Scanning Electron Microscope in Analyzing Rolling Contact Surfaces, Paper 790851, Society of Automotive Engineers, Warrendale, Pa., 1979. Volume II 537 Copyright © 1983 CRC Press LLC GEARS J. L. Radovich INTRODUCTION Many studies have been made in recent years to understand more fully the lubrication requirements of gears. The ideal situation would be a theoretical solution which would predict the optimum lubricant for a specific set of gears and operating conditions based on easily measured system parameters. To date, gear lubrication has not been reduced to this pure science. Consequently, experience is still one of the most valuable tools for proper lubricant selection. Lubrication provides the vital function of separating the contacting surfaces of the gear teeth by an easily sheared film which reduces friction, improves efficiency, and extends the useful life. In addition, lubrication may also provide cooling and flushing of the gear tooth surfaces, corrosion protection, and chemical modification of the surface material. Although proper lubrication is a necessity for successful operation of a set of gears, it is not a cure for inadequate design, manufacture, or improper operation. GEAR TYPES AND TERMINOLOGY Figure 1 shows a spur gear and pinion in mesh and displays several terms used in gearing. Acentral element is the pitch diameter which is calculated as follows: or or where D and d are the pitch diameter of the gear and pinion, respectively; C is the operating center distance; Tand t are the number of teeth in the gear and pinion, respectively; and R is the ratio of the gear set (R = T/t). Pitch line velocity is the peripheral speed of the pitch diameter in meters per second or feet per minute. where v = pitch line velocity in meters per second; d = pinion pitch diameter in meters; and N = pinion speed in RPM; or where v = pitch line velocity in feet per minute; d = pinion pitch diameter in inches; and N = pinion speed in RPM. There are several types of gear configurations (Figure 2). Each gear type has different design advantages and some have special lubrication requirements. Spur — Gear shafts are parallel and gear teeth are cut in line with the shaft centerline (Figure 2a). For spur gears with a transverse contact ratio (contact length divided by base pitch) of less than two, the tangential load is carried by two teeth at the beginning of the Volume II 539 539-564 4/10/06 5:02 PM Page 539 Copyright © 1983 CRC Press LLC contact cycle. The load is then carried by one tooth only as one of the teeth leaves mesh and then by two again as the next tooth comes into mesh. Helical — Gear shafts are usually parallel, but may be at any angle to each other. Gear teeth are cut at an angle to the shaft certerline (Figure 2b). The transfer of load from one tooth to the next is more uniform than spur because several teeth are always in contact along some portion of the tooth face at the same time. Because of the helix angle this gear type generates a thrust load along the axis of the gear shaft. Double helical — Gear shafts are parallel. The gear face is split into two sections, each with helical teeth. The two helical sections have equal helix angles, but opposite hands (Figure 2c). Contact conditions are the same as single helical, but since the thrust load from each helix is equal in magnitude and opposite in direction, no net thrust load is imposed on the gear shaft. However, one of the elements must be free to move axially with respect to the other in order to equalize the tooth loads on each helix. If this is not done, single helix loading will occur. Bevel — Shaft centerlines are orthogonal and intersecting. Bevel gears can be straight or spiral (Figure 2d). Hypoid — Basically the same as bevel gears except that shaft centerlines do not intersect (Figure 2e). Because of this offset, relative sliding velocity between contacting surfaces is higher than for bevels. Because of this sliding and the high contact stresses, an extreme pressure lubricant compounded with friction modifying additives is required. Worm — Shaft centerlines are orthogonal and nonintersecting. The worm resembles a screw thread and drives the worm gear. Both elements are in the same plane (Figure 2f). Since the worm rotates like a screw, high-sliding velocity is developed between contacting surfaces on the worm and wheel. As a result, a lubricant containing friction modifiers is necessary to reduce friction and improve efficiency. DESIGN CONSIDERATIONS AND GEAR MATERIALS In considering a gear application, the power to be transmitted and input speed and gear ratio are usually specified. Orientation of the input shaft to the output shaft may also be indicated. Standard formulas for determining the allowable power which can be transmitted by a gear set have been developed by the American Gear Manufacturers Association (AGMA). Using these formulas in conjunction with the information specified, the designer then has to balance the following variables. Gear type — If the input and output shafts are required to operate at right angles, or some condition other than parallel, bevel, hypoid, helical, or worm gears must be used. If the shafts are parallel to one another, spur, helical, or double helical gears can be used. The gear type will also influence the type of bearings and the housing design required to support the gear forces. Center distance — As the linear distance between the centerlines of two mating gears is increased, for the same transmitted power, the tangential tooth load decreases since the torque is generated with a longer moment arm. The pitch diameters of the gear and pinion would increase and, consequently, the pitch line velocity would increase also. This increase in center distance would allow a narrower face width or softer material and less stringent lubrication requirements. The disadvantages are that the larger gears take more space and tend to cost more. Face width — By increasing the width of the gear face, the contact area is lengthened and unit loading is reduced. This would allow the use of softer gear material or a reduction of the center distance. The disadvantage is that as the face width increases, the shafting must be made more rigid so that dynamic deflection of the gear shaft will not reduce the effective contact of mating teeth. Volume II 541 539-564 4/10/06 5:02 PM Page 541 Copyright © 1983 CRC Press LLC [...]... properly lubricated with this type of lubricant Being so viscous and adherent to gear teeth, this type of lubricant does not offer the advantage either of cooling or flushing the gear mesh Copyright © 1 983 CRC Press LLC 539-564 4/10/06 546 5: 02 PM Page 546 CRC Handbook of Lubrication Table 1 TYPES OF LUBRICANT USED WITH VARIOUS GEAR APPLICATIONS From Root, D C, Lubr Eng., 32, 8, 1976 With permission Greases... approximations of millimeters and degree Celsius shown 5: 02 PM CRC Handbook of Lubrication a 4/10/06 550 Table 4 AGMA LUBRICANT NUMBER RECOMMENDATIONS FOR ENCLOSED CYLINDRICAL AND DOUBLE-ENVELOPING WORM GEAR DRIVES 539-564 e 4/10/06 d 5: 02 PM Worm gears of either type operating at speeds above 24 00 rpm or 10m/sec (20 00 fpm) rubbing speed may require force-feed lubrication In general, a lubricant of lower... toxic or irritating to the skin Consult lubricant supplier’s instructions Viscosities of AGMA lubricant numbers, 13 and above are specified at 21 0 °F ( 98. 9 °C) as measurement of Saybolt viscosities of these heavy lubricants at 100 °F (37.9 °C) would not be practical 5: 02 PM CRC Handbook of Lubrication a 4/10/06 5 52 Table 5 VISCOSITY RANGES FOR AGMA OPEN GEAR LUBRICANTS ... speeds in the temperature range of approximately – 20 to 120 °C This type of lubricant is ideal for bearings if both bearings and gears must be lubricated from the same system Constant relubrication of gear teeth is preferred since the oil does not adhere to tooth surfaces This type of oil can be used effectively to cool the gear mesh and flush the tooth surfaces of wear particles or debris The lubricant... satisfactorily using other types of oils Such oils should be used, however, only upon approval by the manufacturer From Standard AGMA 25 0.04, Lubrication of Industrial Enclosed Gear Drives, American Gear Manufacturers Association, Arlington, Va., 1974 With permission Page 551 Volume II 551 Copyright © 1 983 CRC Press LLC 539-564 From Standard AGMA 25 1. 02 Lubrication of Industrial Open Gearing, American... of insufficient lubrication Of nonferrous gear materials, bronze is the most common Typical bronze alloys for gearing are 86 to 90% copper, 9 to 12% tin, and 3% or less of lead, zinc, and phosphorus This material does not rust, is nonmagnetic, and offers a good balance of strength and hardness It is frequently used for worm wheels which, when run with hardened, ground steel worms, create a system of. ..539-564 4/10/06 5 42 5: 02 PM Page 5 42 CRC Handbook of Lubrication Helix angle — Increasing the angle that is made by the centerline of the tooth with the centerline of the shaft will increase the face contact ratio This means that more teeth are in mesh in the contact zone, which distributes... high-sliding action of the gear teeth requires a friction reducing agent to reduce heat and improve efficiency The useful temperature range is approximately 5 to 120 °C Bearings can be lubricated with this type of lubricant without difficulty Constant relubrication of gear teeth is recommended since this type of oil does not cling to gear teeth and will be wiped off the gear teeth in mesh This type of lubricant... meshing with hardened steel gears They offer quiet operation, abrasion Copyright © 1 983 CRC Press LLC 539-564 4/10/06 5: 02 PM Page 543 Volume II 543 (b) Pitch point (a) First point of contact (c) Last point of contact FIGURE 3 Relative motion of meshing gear tooth surfaces resistance, and resistance to impact loads because of the material resilience Suitability of lubricants depends on compatibility... properties of the lubricant are most important to prevent scoring of the surfaces due to metal-to-metal contact If gear sets were operated under conditions of boundary lubrication for extended periods of time, wear would be rapid and severe With increased relative motion, the gearing moves into mixed lubrication Here, tooth surface asperities are close enough to influence the coefficient of friction . gears. They offer quiet operation, abrasion 5 42 CRC Handbook of Lubrication 539-564 4/10/06 5: 02 PM Page 5 42 Copyright © 1 983 CRC Press LLC resistance, and resistance to impact loads because of the. ASME J. Lubr. Technol., 103 (2) , 29 5, 1 981 . 27 . Neale, M. J., Ed., Tribology Handbook, John Wiley & Sons, New York, 1973. 28 . Szeri, A. Z., Ed., Tribology — Friction, Lubrication, and Wear, Hemisphere. recommendations. 534 CRC Handbook of Lubrication Table 18 (continued) FAILURE MODES THAT LIMIT PERFORMANCE FIGURE 24 . Scanning electron micrograph of surface nucleated spall. Copyright © 1 983 CRC Press LLC Bearings