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,
Trang 2field 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
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Trang 3coating 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 EP or 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
Table 15 TYPICAL OPERATING TEMPERATURE LIMITS FOR
GREASES
Trang 4brication 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.)
Trang 5Table 18 FAILURE MODES THAT LIMIT PERFORMANCE
Trang 6Fatigue 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.
Trang 7Bearings 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 relubrilubri-cation 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
FIGURE 25 SEM of ground surface after running (Magnifications × 100 and 500.)
Trang 81 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.
Trang 9J 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
A central 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; T and 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
Copyright © 1983 CRC Press LLC
Trang 10contact 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
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Copyright © 1983 CRC Press LLC