PART 1

PART 1 Robert Enichello GEARTECH Albany, California Introduction This is a four-part article explaining the principles of gear lubrication. It reviews current knowledge in the field of gear tribology and is intended for both gear designers and gear operators. Part 1 classifies gear tooth failures into five modes and explains the factors that a gear designer and gear operator must consider to avoid gear failures. It defines the nomenclature and gives a list of references for those interested in further research. Pan 2 contains an in-depth discussion of the gear tooth failure modes that are influenced by lubrication. It also gives methods for preventing gear tooth failures. Part 3 gives an equation for calculating the lubricant film thickness which determines whether the gears operate in the boundary, elastohydrodynamic or fullfilm lubrication regime. Also given is an equation for Blok's flash temperature which is used for predicting the risk of scuffing. Part 4 gives recommendations for selecting lubricant type, viscosity and application method. Finally, a case history is given which demonstrates many of the principles of gear lubrication. Gear Tribology Because gears are such common machine components, they may be taken for granted. Not generally appreciated is that they are complex systems requiring knowledge from all the engineering eat treatment manufacturing methods and lubrication meet the requirements of a given

method is as important as the choice of steel alloy and heat treatment. The interrelationship of the following factors must be considered: 1. Gear tooth geometry. 2. Gear tooth motion (kinematics). 3. Gear tooth forces (static and dynamic). 4. Gear tooth material and surface characteristics (physical and chemical). 5. Lubricant characteristics (physical and chemical). 6. Environment characteristics (physical and chemical).

Gear Tooth Failure Modes To obtain optimum, minimum-weight gearsets the gear designer must be aware of the intricate details of many competing modes of failure. The American Gear Manufacturers Association (AGMA) has classified 20 modes of gear failure in their nomenclature publication (I),under the broad categories of wear, surface fatigue, plastic flow, breakage and so-called associated gear failures. See also references (2) through (6) for gear failure modes. For our purposes, the best choice for the basic categories is as follows: 1. Overload. 2. Bending Fatigue. 3. Hertzian Fatigue. 4. Wear. 5. Scuffing. The following lists subdivide the five basic failure modes. Many gear failures are known by several names and/or qualifying terms such as: initial, moderate, destructive, etc. These names and terms are included in the lists in parentheses. The term "scoring" has been used in the past in the U.S.A., while the term "scuffing" is used in Europe to describe the severe form of adhesive wear which involves welding and tearing of the surfaces of gear teeth. To agree with current usage, the term scuffing will be used in this article when referring to this failure mode. The term scoring implies scratching and it will be used to describe abrasive wear rather than scuffing. 1 - Overload Brittle fracture. Ductile fracture. Plastic deformation. cold flow hot flow (continued on next page)

indentation (rolling, bruising, peening, brinelling) rippling (fish scaling) ridging bending, yielding tip-to-root interference

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2 Bending Fatigue Low-cycle fatigue ( G 1000 cycles to failure). High-cycle fatigue (> 1000 cycles to failure).

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3 Hertzian Fatigue Pitting (initial, superficial, destructive, spalling). Micropitting (frosting, grey staining, peeling). Sub-case fatigue (case crushing).

4 - Wear Adhesion (normal, running-in, mild, moderate, severe, excessive). Abrasion (scoring, scratching, plowing, cutting, gouging). Corrosion. Fretting-corrosion. Cavitation. Electrical discharge damage. Polishing (burnishing).

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5 Scuffing Scuffing (scoring, galling, seizing, welding, smearing, initial, moderate, destructive).

Symbols and Units

Symbol BM

bH c CM

d El,& Er

hi,

Description thermal contact coefficient semi-width of Hertzian contact band constant specific heat per unit mass operating pitch diameter of pinion modulus of elasticity (pinion, gear) reduced modulus of elasticity minimum film thickness minimum contact length transmitted power oil flow rate average surface roughness, rms bulk temperature

Units lbf/[in so,' OF] in h~/g~m lbf in/@bOF] in

Tb, test Tc Tf Tf, test

Ts Ve Vr,, Vr2 V WN~ W N ~

XW Ibf/in2 Ibf/in2 in in v" hp gPm pin

OF

x, u A AM

Pm pa "I

,"2

"40

P I ,P2 P~ pn u

QI,~, Wb

o,,w2

bulk temperature of test gears contact temperature flash temperature maximum flash temperature of test gears scuffing temperature entraining velocity rolling velocity (pinion, gear) operating pitch line velocity normal operating load normal unit load welding factor load sharing factor pressure-viscosity coefficient specific film thickness heat conductivity mean coefficient of friction absolute viscosity Poisson's ratio (pinion, gear) kinematic viscosity at 40 OC transverse radius of curvature (pinion, gear) density normal relative radius of cunature composite surface roughness, rms surface roughness, rms (pinion, gear) base helix angle angular velocity (pinion, gear)

OF OF OF OF OF in/s in/s ftlmin Ibf lbflin

in2/lb lbf/[s OF] Reyns (Ib s/in2) cSt in Iblin3 in pin pin deg rad/s

References 1 ) "Nomenclature of Gear Tooth Failure Modes", ANSIlAGMA 110.04, (1980). 2) Shipley, E. E., "Gear Failures," Machine Design, pp 152-162, Dec. 7, (1967). 3) Dudley, D. W., "Gear Wear," Wear Control Handbook, ASME. 4) Ku, P. M., "Gear Failure Modes - Importance of Lubrication and Mechanics," STLE Trans., 19, No.3, pp 239-249 (1975). 5) Wulpi, D. J., "How Components Fail." 6) "Failure Analysis and Prevention," (Failures of Gears), Metah Handbook, 10,8th ed., pp 507-524. 7) God:frey, D., "Recognition and Solution of Some Con~monWear Problems Related to Lubrication and :Hydraulic Fluids," Lubrication Engineering, pp 111-114, Feb. (1987). 8) Littman, W. E., "The Mechanism of Contact Fatigue," Interdisciplinary Approach to the Lubrication of Concentrated Contacts, NASA SP-237, pp 309-377 (1970). 9) Ueno, T., et. al., "Surface Durability of CaseCarburized Gears - On a Phenomenon of "Grey - ctsining of Tooth Surface," ASME Pap. NO. :2/DET-27, pp 1-8 (1980). ter, H. and Weiss, T., "Some Factors ~encingthe Pitting, Micropitting (Frosted AreeIS)and Slow Speed Wear of Surface Hardened Gerurs," ASME pap. No. 80-C2/DET-89, pp 1-7 (i98r4.

11) Shipley, E. E., "Failure Analysis of Coarse-Pitch, Hardened and Ground Gears," AGMA Pap. No. P229.26, pp 1-24 (1982). 12) Tanaka, S., et. al., "Appreciable Increases in Surface Durability of Gear Pairs with Mirror-Like Finish," ASME Pap. No. 84-DET-223, pp 1-8 (1984). 13) Adams, J. H. and Godfrey, D., "Borate Gear Lubricants-EP Film Analysis and Performance," Lubrication Engineering, Vol. 37, No. 1, pp 16-21, Jan. (1981). 14) Blok, H., "Les Temperatures de Surface dans Les Conditions de Graissage Sons Pression Extreme," Second World Petroleum Congress, Paris, June (1937). 15) Blok, H., "The Postulate About the Constancy of Scoring Temperature," Interdisciplinary Approach to the Lubrication of Concentmted Contacts,NASA SP-237, pp 153-248 (1970). 16) Dowson, D., "Elastohydrodynamics," Paper No. 10, Proc. Inst. Mech. Engrs., Vol. 182, FT 3A, pp 151-167 (1967). 17) "Fundamental Rating Factors and Calculation Methods for Involute Spur and Helical Gear Teeth," AGMA 20014388, (1988). 18) Akazawa, M., Tejima, T. and Narita, T., "Full Scale Test of High Speed, High Powered Gear Unit - Helical Gears of 25,000 PS at 200 mls PLV," ASME Pap. No. 80-C2/DET4 (1980). 19) Drago, R. J., "Comparative Load Capacity Evaluation of CBN-Finished Gears," AGMA Pap. No. 88 FTM 8, Oct. (1988). 20) "AGMA Standard Specification - Lubrication of Industrial Enclosed Gear Drives," AGMA 250.04, Sept. (1981). 21) "Practice for High Speed Helical and Herringbone Gear Units," AGMA 421.06, Jan. (1969). 22) WeIlauer, E. J., and Holloway, G. A., "Application of EHD Oil Film Theory to Industrial Gear Drives," T m . ASME, J. Eng. I d , Vol. 98, series B., No. 2, pp 626-634, May (1976). 23) Drago, R. J., "Fundamentals of Gear Design," Butterworth, (1988).

TMX-77572,Oct. (1974). 25) SCORING +, Computer Program, GEARTE Software, Inc., Copyright 1985-1989.

PART 2 BY Robert Errichello GEARTECH Albany, California

.%tor's Nore: n e ~ l l o w m gI S h n 2 of a 4-pan amcle. h n I appeared rn the January 1990 rssue of f.E h n s 3 and 4 wtll be publrshed m rhe March and Apnl Issues respecfrvely.

Introduction This is a four-part article explaining the principles of gear lubrication. It reviews current knowledge of the field of gear tribology and is intended for both gear designers and gear operators. Part 1 classifies gear tooth failures into five modes and explains the factors that a gear designer and gear operator must consider to avoid gear failures. It defines the nomenclature and gives a list of references for those interested in further research. Part 2 contains an in-depth discussion of the gear tooth failure modes that are influenced by lubrication. It also gives methods for preventing gear tooth failures. Part 3 gives an equation for calculating the lubricant film thickness which determines whether the gears operate in the boundary, elastohydrodynamic or full-film lubrication regime. Also given is an equation for Blok's flash temperature which is used for predicting the risk of scuffing. Part 4 gives recommendations for selecting lubricant type, viscosity and application method. Finally, a case history is given which demonstrates many of the principles of gear lubrication.

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Lubrication Related Failure Modes These articles are concerned with gear tooth failures that are influenced by lubrication. Pitting or scuffing may cause the gear teeth to deteriorate and generate dynamic forces which in turn cause the gear teeth to fail by bending fatigue. In these cases the bending failure is secondary, and not directly related to lubrication, while pitting or scuffing are

the primary failure modes and both are definitely influenced by lubrication. The failure analyst must discern the difference between primary and secondary failure modes because the wrong corrective action is likely to be recommended if a secondary failure mode is mistaken for the primary failure mode. For example, increasing the size of the gear teeth to prevent reoccurrence of the above mentioned bending failure would only make the situation worse by lowering the pitting and scuffing resistance. Godfrey (7) gives a good description of lubrication-related failure modes. With the above considerations, overload and bending fatigue are judged to be unrelated to lubrication and are eliminated from further discussion together with sub-case, Hertzian fatigue. Although corrosion, frettingcorrosion, cavitation and electrical discharge damage are influenced by lubrication, they are not discussed because these failure modes occur relatively rarely in gear teeth. Hence, the following failure modes are included in the scope of this article: Hertzian fatigue pitting micropitting Wear adhesion abrasion polishing Scuffing

Hertzian Fatigue Pitting is a common failure mode for gear teeth because they are subjected to high Hertzian contact stresses and many stress cycles. For example, through-hardened gears are typically designed to withstand contact stresses of approximately 100,000 psi while the contact stresses on carburized gears may reach 300,000 psi. In addition, a given

tooth on a pinion that is revolving at 3600 rpm accumulates over 5 million stress cycles every 24 hours. Pitting is a fatigue phenomenon (8) which occurs when a fatigue crack initiates either at the surface of the gear tooth or at a small depth below the surface. The crack usually propagates for a short distance in a direction roughly parallel to the tooth surface before turning or branching to the surface. When the crac:ks have grown to the extent that they separate a piece of the !surface material, a pit is formed. If several pits grow togelther to form a larger pit it is often referred to as a "spall." Ther-e is no endurance limit for Hertzian fatigue, and pitting OCCUIrs even a1t low stresses if the gears are operated long enou~gh.BecauIse there is no endurance limit, gear teeth must be designed fc)r a suitable, finite lifetime. .. l o extena tne pitting life of a gearset, the designer must keep the contact stress low, material strength high and the lubricant specific film thickness high. There are several geometric variables such as diameter, face width, number of teeth, pressure angle, etc., that may be optimized to lower the contact stress. Material alloys and heat treatment are selected to obtain hard tooth surfaces with high strength. Maximum pitting resistance is obtained with carburized gear teeth because they have hard surfaces, and carburizing induces beneficial compressive residual stresses which effectively lower the load stresses. The drawback is that they are relatively expensive to produce because they must be finished by grinding. The details for obtaining high lubricant specific film thickness will be explained later when elastohvdrodvnamic (EHD) lubrication is discussed, but general recommendations are to use an adequate supply of cool,, clean anc1 dry lubricant that has adequate viscosity and a hit;h pressure-viscosity coefficient. Pilung 111ayinitiate at the surface or at a subsurface defect such as a nonmetallic inclusion. With gear teeth, pits are most often of the surface-initiated type because the lubricant film thickness is usually low, resulting in relatively high metalto-mletal contact. The inte:raction between asperities or contact at defects, suck1 as nicks or furrows, creates surface-initiated craclcs rather . . tllan subsurface-initiated cracks. For high-speed gears wit11 smootn surhce finishes, the film thickness is greater and sub-surface initiated pitting may predominate rather than surface-initiated. In these cases, pitting usually starts at a subsurface inclusion which acts as a point of stress concentration. Cleaner steels, such as those produced by vacuurn melting, prolong the pitting life by reducing the numlber of inclusions. Ccbntamination from water in the lubricant is believed to promote pitting through hydrogen embrittlement of the metal, and abrasive particles in the lubricant cause pitting by indenting the tooth surfaces causing stress concentrations ---'IT disruding the lubricant film. At present, the influence bricant' ac~ditivesorI pitting is unresolved.

.

-

..:--

!

I

Metlhods For Preventin~gPitting 1. Reduce contact stresses by reducing loads or optimizing gear geometry. 2. Use clean steel, properly heat treated to high hardness, :eferably by carburizing.

3. Use smooth tooth surfaces produced by careful grinding or honing. 4. Use an adequate amount of cool, clean and dry lubricant of adequate viscosity.

Hertzian Fatigue - Micropitting On relatively soft gear tooth surfaces, such as those of through-hardened gears, Hertzian fatigue forms large pits with dimensions on the order of millimeters. With surface hardened gears such as carburized, nitrided, inductionhardened and flame hardened, pitting may occur on a much smaller scale, typically only 10 r m deep. To the naked eye, the areas where micropitting has occurred appear frosted, and "frosting" is a popular term for micropitting. Japanese researchers (9) have referred to the failure mode a s "grey staining" because the light-scattering properties of micropitting gives the gear teeth a grey appearance. Under the scanning electron microscope (SEM) immediately evident is that micropitting proceeds by the same fatigue process as classical pitting, except the pits are extremely small. Many times micropitting is not destructive to the gear tooth surface. It sometimes occurs only in patches, and may arrest after the tribological conditions have improved by running-in. The micropits may actually be removed by light polishing wear during running-in, in which case the micropitting is said to "heal." However, there have been examples (9), (10)and (11) where rnicropitting has escalated into full scale pitting, leading to the destruction of the gear teeth. The specific film thickness is the most important parameter that influences micropitting. Damage seems to occur most readily on gear teeth with rough surfaces especially when they are lubricated with low viscosity lubricants. Gears finished with special grinding wheels to a mirror-like finish (12) have effectively eliminated micropitting. Slow-speed gears are prone to micropitting because their film thickness is low. To prevent micropitting, the specific film thickness should be maximized by using smooth gear tooth surfaces, high viscosity lubricants and high speeds. Experiments (10)have shown that flame-hardened and induction-hardened gears have less resistance to micropitting than carburized gears of the same hardness. This is probably due to the lower carbon content of the surface lavers of the flame-hardened and induction-hardened gears. Methods For Preventing Micropitting 1. Use smooth tooth surfaces produced by careful grinding or honing. 2. Use an adequate amount of cool, clean and dry lubricant of the highest viscosity permissible. 3. Use high speeds if possible. 4 . Use carburized steel with proper carbon content in the surface layers.

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Wear Adhesion Adhesive wear is classified as "mild" if it is confine"a 10 the oxide layers of the gear tooth surfaces. If, however, the oxide layers are disrupted and bare metal is exposed, the transition to severe adhesive wear usually occurs. Seivere adhesive wear is termed scuffing and will be discussed liiter. For the present, we assume that scuffing has been avoi;ded through proper design of the gears, selection of the lubric:ant and control of the running-in process. When new gear units are first operated the contact betu the gear teeth is not optimum because of unavoid;_._ manufacturing inaccuracies. If the tribological conditions are favorable, mild adhesive wear occurs during running-in and A -

l ENGINEERING

r subsides with time, resulting in a satisfactory lifetime for the gears. The wear that occurs during running-in is beneficial if it smoothes the tooth surfaces, increasing the specific film thickness, and increases the area of contact by removing minor imperfections through local wear. To ensure that the wear rate remains under control, run-in new gearsets by operating for at least the first 10 hours at one-half load. The amount of wear considered tolerable depends on the expected lifetime for the gears and requirements for control of noise and vibration. Wear is considered excessive when the tooth profiles wear to the extent that high dynamic loads .\rellr or the tooth thickness is reduced to the extent that bending fatigue becomes possible. because of practical limits on lubricant Man.y gears, viscosity, speed and temperature, must operate under boundary-lubricated conditions where some wear is inevitable. Highly-loaded, slow speed ( 100 fpm), boundary-lubricated gears are especially prone to excessive wear. Tests with slow-speed gears (10) have shown that nitrided gears have good wear resistance while carburized and through-hardened gears have similar but lower wear resistance.-Reference (10) concluded that lubricant viscosity has the greatest influence on slow-speed, adhesive wear and that high viscosity lubricants reduce the wear rate significantly. Also found was that sulphur-phosphorous additives can be detrimental with slow-speed ( 10 fpm) gears, giving very high wear rates. A few gear units operate under ideal conditions with smooth tooth surfaces, high pitch line speed and high lubricant film th;J--sss. For example, turbine gears that operated almost continr~ouslyat 30,000 fpm pitch line speed still had the origina1machining marks on their teeth even after operating for 20 years. Most gears however, operate between the bound:3ry and full-film lubrication regimes, under elastohydrodynamic (EHD) conditions. In the EHD regime, with thle proper type and viscosity of lubricant, the wear rate usuallv reduces during running-in and adhesive wear virtually ceases once running-in is completed. If the lubricant is proper1y maintained (cool, clean and dry) the gearset should not suf 'fer an adhesive wear failure.

A.."""




U

yns (lb slec/in2) PO = abs~olutevisc viscosity versus Fig. 3 gives average. v. temperature for typlcal nunem gear lubricants wit1i a viscosity index

..

a

- pres:sure-viscosity coefiScient, (in2/lb). . . "- .

The pressure-viscosity coerrlclent fanges from a = 0.5x10-4to a = Zx104 in2/lb for typical gear lubricants. Data for pressure-viscosity coefficients versus temmrature for typical gear lubricants are givc:n in Fig. 4.

'"

20

>

e

42

2 10 8

1 .o 8

0.8 0.8 80

100

120

140

160

180

200

4

220

BULK TEMPERATURE. (OF)

FIG. 3 - Absolute viscosity versus temperature for mineral gear lub~ with a viscosity index of 95 (17). BULK TEMPERATURE. (OC)

= reduced modulu!3 of elasticity given by:

where isson's ratio (pinion, gear). odulus of elasticity (pinion, gear). Pn = norm~ a lrelativc:radius of curvature.

pinion, ge,ar). = base

. ..

le. 250 50 1!OO ULK TEMPER.ATURE, (OF)

"'50

Ve

=

entraining velacity givenI by:

V* = Vrl

+ Vr,

t m

FIG.4 .

300

~scositycoef ficient versu:s temperatulx. for minerad gear (17).

5 Fi 9

Xp

E

UNMODIFIED TOOTH PROFILES

1 -

0

2 ,-

Load ollal lllB Factor, Xr The load sharing factor accounts for load sharing between succeeding pairs of teeth as influenced by profile modification (tip andlor root relief) and whether the pinion or gear is the driver. Figure 5 gives plots of the load sharing factors for unmodified and modified tooth profiles. As shown by the exponents in the Dowson and Higginson equation, the film thickness is essentially determined by the entraining velocity, lubricant viscosity and pressure-viscosity coefficient, while the elastic properties of the gear teeth and the load have relatively small influences. In effect, the relatively high stiffness of the oil film makes it insensitive to load, and an increase in load simply increases the elastic deformation of the tooth surfaces and widens the contact area, rather than decreasing the film thickness.

4

9 I

0,

€2

€1

€5

€4

PINION ROLL ANGLE

D TOOTH PROFILES PINION DRIVING

GEAR DRIVING

Blok's Contact Temperature Blok's (14) contact temperature theory states that scuffing will occur in gear teeth that are sliding under boundarylubricated conditions, when the maximum contact temperature of the gear teeth reaches a critical magnitude. The contact temperature is the sum of two components, the bulk temperature and the flash temperature, i.e.,

€2

€5

€4

PlNlON ROLL ANGLE

FIG. 5 - Load sharing factor versus pinion roll angle (17).

Blok's flash temperature equation as formulated in AGMA 2001-B88, Appendix A (17) for spur and helical gears is:

where (

5o

)

50 - S S

where

=

3.0

average surface roughness, rms.

pm = mean coefficient of friction. X, = load sharing factor. W N ~=

normal unit load.

Thermal Contact Coefficient, BM The thermal contact coefficient is given by:

Vrl = rolling velocity of the pinion. Va = rolling velocity of the gear. BM= thermal contact coefficient. bH = semi-width of Hertzian contact band.

Mean Coefficient of Friction, pm The following equation gives a typical value of 0.06