Gear Lubrication 3

7/27/2019 Gear Lubrication 3

1/91 , 4 GEAR TECHNOLOGY

Th e L u b r i c a t io n o fGea r s - Pa r t III

R o b e rt E n ic h el l! oG i E A R T I E C H . Albany , . le A

IntroductionThis is the final part of a three-part series on

the basics ofgear lubrication. It covers selectionof lubricant types and viscosities. the applicationof lubricants. and a case history.

Selecting Lubricant. TypeThe choice of lubricant depend onthe type

of gearing and enclo ure, operating speed andload. ambient temperature, and method oflubri-cant application. Most gears are lubricated withone of the following types: oil. grease, adhesiveopen-gear lubricant, Dr solid lubricant. The opti-mum lubricant for any application is the leastexpensive, considering both initial cost and main-tenance co ts, thai meet the requirements.

Oil is ihe rna 'Iwidely used lubricant becauseit is readily distributed to gears and bearings andhas both good lubricating and cooling propertie .Also. contamination may be readily removed byfiltering or draining and replacing the oil. How-ever, it requires an oil-tight enclosure providedwith adequate haft eals.

Grea e is suitable only for low-speed, low-'load applications because il.does not circulatewell, and it is a relatively poor coolant. Greaselubricated gears are generally boundary lubri-cared becausethe grease is either pushed asideor thrown from the gear teeth, Contaminationfrom wear particles or other debris is u uallytrapped in the grea eend require COSIlymaintenance to-eliminate. Greaseis often usedto avoid leakage from enclosures that are noton-light. However .. if all the factors are con-sidered, it is usually found that an oil lubricantis more economical and reliable than a greasefor gear lubrication.

Open-gear lubricants are viscous, adhesivesemi-fluids used on large. low-speed, open gears.

such as those used in iron ore and cement mills,antenna drives. bridge drives. cranes, etc. 'Gearsin the e applications run slowly, and they aretherefore boundary lubricated, The lubricantmu t bond strongly to resi t being thrown off thegear teeth. However. the queezing and lidingaction of gearleeLh tends 10push the lubricant into'the root of the gear teeth where it is relativelyineffective. These iubricants are applied by handbrushing or by automatic systems which d sliveran intermi t tent spray ..Some open gear lubricantsare thinned with II quick-evaporating olvent/diluent to make them easier to apply, Open-gearlubricants share the disadvantages of grease lu-brication, and they are especially costly (andmessy) to maintain. For these reasons. the 'trendi away from open gear toward enclo ed, oil-lubricated gearboxes whenever possible,Solid Ilubricants, usua\Jy in the form ofbonded,

dry films. are 1I ed where the temperature.i toohigh ortoo [ow for an oil orgrea e: where leak agecannot be tolerated; or where the gears mustoperate ina vacuum. These lubricantsare usuallymolybdenum disulfide (MoS2) or graphite in aninorganic binder, which is appliedto the gearteeth and cured 10 form a dry film coating.Polytetrafluoroethylene (PTFE) and tungstendisulfide (WS2 ) coating are also used. Solidlubricants are expensive to apply and have limitedwear lives. However. in many applications, suchasspacec raft, they are the only alie mati ve and canprovideexcelleut service.

Only oil lubricants will be discus ed in greaterdetail. Gil should be used as the lubricant unlesthe operating conditions preclude its use. Gener-ally. the simplest and least expensive lubricationsystem for gears is a totally enclosed. oil-bathof mineral oil.


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The lubrication requirements of pur. hell-cat, straight-bevel. and spiral-bevel gearsareessentially the same. For thi clas of gears, themagnitudes of the loads and .Iiding peeds aresimilar, and requirement .forvi cosity and anti-scuffpropertie are virtually identical. Manyindu trial spur and helical. gear units are lubn-caeed with [U t and oxidation-inhibited (R&O)mineral oils, The low viscosity R&O oils.commonly called turbine ,oils, are used in manyhigh- peed gear unit, where the gear [Oathloads are relatively low. Mlneral oll withoutanti- cuff additives are uitable forhigh-speed,hghtly loaded gear where the high entrainingvelocity of the gearteeth developsthick EHDoil films. In these CO l e the mo t importantproperty of the lubricant i vi co ity, Anti-scuff/EP additives are unnecessary because thegear teeth are separated, elirninatiag metal-to-metal contact and the cuffing mode of failure.Slower speed gears, especially carbu rized gears,tendto be more heavily loaded. The e gearsgenerally require higher viscosity lubricantswith anti-scuff additives.

Hypoid gear. such as those used for auto-motive axles, are especially prone to scuffingbecause they are heavily loaded and they havehigh sliding velocities, For these reasons,hypoid gear oils have the higher concentrationsof anti-scuff additive.For critical applications ..the contact tempera-tore hould be calculated with Blok' Iequationand compared to the scuffing temperature of thelubricant. Thi quantitative method is effectivefor selecting a lubricant with adequate cuffingre sis ta nc e .

Worm gears have high hdingvel.ocitywhich generates s.ignificant frictional losses.Fortunately. their tooth load are relativelylight. and they are successfully lubricated withmineral oil. that are compounded with lubric-ity additives. These oils eontain 3% to 10%fany oil or low acid tallow, The polar mol-ecule of the additive form urface films byphy ical ad orption or by reaction with thesurface oxide to form a metallic soap whichacts as a [ow hear strength film. improving the"lubricity" or friction-reducing properly.

Synthetic lubricants are u ed for appJ ica-tions, such as aircraft gas turbines, where the oilmust operate over a wide temperature range andhave good oxidation stability at high tempera-

ture. Ester and bydrocarbon synthetic lubricantshave high viscosity indices. giving them goodfluidity or low vi cosities at very low tempera-tures and acceptable viscositie at high tempera-tures. The volatility of e ters is lewerthno thatof mineral oils ofthe same vi cosity, thus reduc-ing oil loss at high temperature. Despite theirlong service life. the extra co t of syntheticlubricants generally cannot be ju tified for oil-bath ystems unless there are extreme tempera-tures involved. becau e the oil mu l be changedfrequently to remove contamination.

Selecting Gear Lubricant Vi co i t . yThe recommendations of AGMA 250.042

~hould be followed when electing lubricantsfor enclosed gear drives that operate at pitchline velocities up to 5,000 fpm. AGMA 421.063should be consulted for high-peed drives (>5,000 fpm).

In our discussion of gear failure' modes, wefound that viscosity is one of the most importantlubricant properties. and the higher the viscos-ity. the greaterthe protection against the variousgeartooth failures. However. the viscosity mustbe limited to avoid excessive heat generationand power loss from churning and shearing ofthe lubricant by high-speed gears or bearings.The operating temperature of the gear drivedeterrmnes the operating viscosity of the lubri-cant. U the lubricant is too vi cou , exce siveheal i generated. T.he heal rai e the lubri-cant temperature aad reduces its viscosity,reaching a point ofdrrnini hing returns whereincreasing the starting viscosity of the lu-bricant leads to 3 . higher operating tempera-ture and a higher oxidation rate, without asignificant gain in operating viscosity,

Gear drives operating ill cold climates must11ve a lubricant that circulate s freely and does notcause high starting torques. A candidate gearlubricant hould have a pour point at Jea t 5C(9f) lower than the expected minimum ambientstart-up temperature, Typical pour points :formineral gear oils are lOaF while synthetic gearlubricantshave significantly lower pour point ofabout -40QF, Pour point depre ants are u ed 1 1 0ta.ilor pour points of mineral lubricants for auto-:motive hypoid gears to. be as low as -40F.

The pitch line speed of the gear is a goodindex of the required viscosity. An empirical.equation for determining required viscosity is

R o b ed E rr ic h eU ois lire p rin cip o; ;1 1GEARTECI:{ . a gear( 'onsul 'r i ll .q fi rm in lbull.l'.CA . H is a rti cle r ep ril lle dhere lias wall the STU's1990 Wilber DeutcnMemoria l Award [or Iheb est a nicle o n th epractiral aspect. 01lubrication. Mr.Errichello is a member ofASME. AGMA. and is (IReg is te r ed Pro fe s si ona lE ng in ee r i n till' S ta te o fCalifornia.JULV/AUGUST 19S' 1 5


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,,1) up' by pitting fatigue. The empirical. equation for7000 f hi 's a pp li ca ti on g i v e

(2 )V40::-----(V) O .S 7000v40 ::: ::: 128 cS t,(3000) 0.5wherev 40 = lubricant kinematic viscosity at40C, cSt

V := o p er at in gph ch l in e velocity, ft/minv= 0 .2 62 d nd = operating pitch diameter of pinion, in .n= pinion speed, rpmCaution must. be 'u ed when using AGMA

recommendation forviscosity. Theauthorknowsof an application where two gear drives wereconsidered to be high-speed. The pinion speedwas 3.625 rpm, qualifying th e gear units as high-speed gear drivesper AGMA 421.06. The geardrives were supplied with oil having the recom-mended viscosity per AGMA 421.06 of ISO 68.However, because the plnion was relatively small,its pitch line velocity was only 3,000 fpm. Thisqualifies the geardrives as slow-speed per AGMA250.04, which recommends a viscosity of ISO150. Both gear dri ves failed within weeks of start-

This indicates that the viscosity per AGMA42 L06 (68 cSt) is much too low, and the visco ityper AGMA 250.04 050 cSt) is appropriate.Hence, definitions of high- peed ver us low-speed must be carefully con idered, and pitchline velocity is generally abetter index thanshaft peed. Thegear drives were rebuilt. withnew gearsets and the ISO VG 68 oil was replacedwith ISO VG 150. The gear drives now operatewithout overheating. a nd the pitting ha b eeneliminated.

For critical applicatinns, the specific filmthickness should be calculated with Dow 011 an dHigginson's" equation. The specific film thick-ness Ii a u efulmeasure of the lubrication reg ime.It can be used with Fig. ~as an approximate guideto the probability of wear-rei ated surface distress.

3 . 02.'0

,," ' " 1.0 iIen ' I ,tn1 . . 0 . 1 I:z : 10.5 ,~ 1 , 1~

I Ii= I' Ii '1: : : l i :. . . . . , 0 ..21::.w ,1 . . 1 . . 01 .1 11 . . 0 . 1n,U'l

I10.05 V '" "

I I II 5%I 40% I I J

, . a D % .,

,I : , II I I I : ' 1

II ' II ,,

,. , ., ,

I I___ II,I 1 0 0 ' l , . . . . . . . I I

ht'Tttt__ - t I ~ " " " 1 l l i - l i !JIll', "~+I. . f . " , , 7 o t = = - f " '_ ~~""-'~.""'-i-.l-PROBABIUTY OF WEA~ 1-'.I' ~ .~ , :RELA .TeD ' !C lSTRESS 1vv ~~I i' ' II / V ,I I II '

I II II 1 1,./1 ,

I I[I 'I I I

,I

I

,

II I,

I

1 0 0 0 0 5 0 0 0 0

1/ II I I ,, I'

I I

0.02 I I I ;0.01 50

1

II I

100 500, 1000 5000PITCH UNE VELOCITY. FPM

Fig. 1- Probability of wear distress, percent. (Fmm AGMA 2001-888, 1988.)

1 6 GEAFI TECH",OlOa'f


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Fig, 1 i based on the data of Wellauer andHolloway,S which were obtained from severalhundred l ab or ato ry t es ts a nd f ie ld a pp li ca ti on s o fgear drives.

Applying, 'Gear Lubr icantsThe method o f applying the lubricant to the

g ear tee th de pe nds fo r th e mo. t pari. primarily onthe pitch line velocity,

Splash lubrication y tern are tile irn-

For very high-speed gears" (above ]6,10100fpm), there is a danger that the amount of oilcarried to the incoming side of the gear me hmay be inadequate. a n d it is prudent to a dd aupplementary flow at 'IIIeincoming side of thegear mesh. Generally, about 2/3 of the oilflow should be supplied to the outgoing sideof the mesh for cooling. and 113 of the flowdirected at the incoming ide for lubrice-tion. The placement of the oil jets is a

plest, but Ihey are limited to a pitch line veloc- crucial factor when pitch line velocitiesex-ity ofabout 3.000 fpm. T,IIe gears should dip ceed 20.000 fpm. At peeds this high. ex-into the oil bath for about twice the tooth depthto provide adequate splash for p inions andbearings and to reduce losses due to churning.The gear hou ing should have troughs to cap-ture the oil flowing down the housing walls.channeling it to the bearings.

periment are required to find the optimumnumber and location for the oil jets.

In pressure-fed ysterns, the followingpa-rameten must be COli idered (Q ensure adequatelubrication a n d cooling of the gear mesh: Qua n -lity offlow.jet size. feed pressure, a nd number -o f

The range of splash lubrication can b e ex- jets. There are general guidelines, based on expe-tended to about 5 , 0 0 0 fp m by using baffles a nd o ilpans te reduce churning. However. above 3.000fprn, providing auxiliary cooling with fans andimproving beat t ransfer by adding Ifill. to thehousing is usually necessary.

Above 5.000 fpm, most gears are lubricatedby a pressure-fed syslem. Fo r gearboxes withan ti fr ic ti on bear ings, sprayingrhe oil al the gearmesh only and re.lying all pia h to lubricate thebearings is perm is ib le up to a pitch line velocityof 7 i O O O [p m m a xim u m. Above this s p e e d , and:for gear drives with journal bearings, both thegears and bearings should be pre ure-fed.

Tile oil.jet should b e placed on the incomingside of thegear mesh for pitch line velocities upto 8,000 fpm. Above 8.000 rpm. more oil isneededforceolingthan for lubricating. and the oilflow removes heat best by being directed at. theoutgoing ide of the gear me h where the oil jetscan trike the hoI, drive-s ide of the gear teeth,

rience and experimentation, for specifying theseparameters, but each application m il t be evalu-ated independently ba ed on it particularoperat-ing conditions and requ i remen t s,

All,empirical equation used to calculate thequantity of oil flow in gallons per minute is

q ; ; ; ; ; PIcwhere c is taken from Table t

p= transmitted power. hpq = oil flow rate, gpm

For a typical indu trial application transmit-ling 2 hp, where weight isnot critical, thede igner might choose the constant c = 200hplgpm, resulting in a copious flow of ] gpm. Onthe other hand, for a high efficiency aviationapplication transmitting 200hp. where weight lscritical, c = 800 might b e chosen. re ulting in alean flow of 0'.25 gpm, Some applications mayrequire different flow rate than tho e given by_able I. . For instance, wide -f ac e. h igh -s pe e d

Table 1 ..Typical. . o n !F lo ws Per 'G ear M eshc(hp/gpmj2004008001000

FlowConditionsCopiousAdequateLeanStarved

CommentGeneral indu trialTyp ic al a v ia ti onL ig ht -w e ig h t, h ig h -e f fi ci ency av ia ti onOnly for unusualconditions

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provide complete Iubricationcoverage of tile facewidth. More than. one jet for each gear me h isadvisable because of th e p ossib ility o f clogging ..The upper l imit I O n t he number of jets is deter-mined by the flow rate and jet diameter; too manyjets for a given flow rate will resultin a jetdiameter less than the minimum. recommended.

gearing may require a higher flow rate toensureuniform cooling and full-face coverage.

The proper jet size, feed pressure ..and num-ber of jets must. be determined to maintain theproper flow rate.jetvelociry, and full-facecover-age. The diameter of a jet can be calculated for a .given flow rate and pressure ba ed on the viscos-ity of the oil'at the operating temperature. 7 Thereare practical limitations onjet: ize, and the mini-mum recommended size is 0.03". U ajet smallerthan Ibis is used , con tami ll la i li s in th e os l may clogit. Typicaljetdiarneters range framO.03"- 0.12".

The feed pressure determines the jet. velocity,which in tum determines the amount of oil thaipenetrates the gear mesh. Typical feed pressuresrange from 20.-WOpsig. lndustrialapplicaticnfeed pressures are typically 3 D psig, and high-speed aerospace applications. are typically 100psig. hl general, the higner the pressure. thegreater the cooling.,8 but the higher the pressure,the smaller the jet diameter. Therefore, pressure:is limited by the minimum recommended jetdiameter of 01.0.3".

The number of jets should be sufficient to

Case Histo ryI l l . an lndustrialapplieation, 24 speed-in-

creaser gearboxes were used to traasrnit 346horsepower and increase speed from 55 rpm 10375 rpm. The gears were parallel-shaft, singlehelical, carbunzed, an d g ro un d. T he sp lasb lubri-cation system used a mineral oil without anti-scuff additives with ISO H>Oviscosity. Afterabout 250 . hours of operation. two gearboxesfailed by bending fatigue ..The gear tooth profileswere so badly wom determining th e primaryfailure mode was impossible .. Three other gear-boxes with less service were elected for in pee-lion. One had logged] 5 hours, while the otherIWO had operated for 65 hourseach. Upon disas-. embly, no broken teeth were found, but all three

30 40 s o 100 110 12 0 ,20 0100SO

1 6 04030~. . . . . 2 0,:f~, 100E 8

,..: 6I~Ui 40u.Ul.s~ 2:l...J! 7 !!i

1 1 . 0o . a0 . 6 ,6

" ' t - - . . . ~ " . . . . . . . . . . . . . . . . 1 1i ' . . . . "i"', ~


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BU LK T EMPERA TURE , (0C)';0 20 301 40 SO 60 7080 90 100 no 120 130140 1501 1:60 170

I,I

gearb oxe . h ad scuffed ge ar te eth . T he prim aryfailu re m o de w as sc uffin g, a nd th e e arlie r b en d-in g f atig u e f ailu re s w e re caused b y d yn am ic lo adgenerated by the w orn gear tee th . Subsequentin p ectio n o f th e re m ain in g g ea rb ox es re ve ale dth at a li h ad s cu ff in g d am a ge , w h ic h p ro ba bly h ado cc urre d im m e diate ly u po n start-u p b eca use th elo ad s w ere n ot re du ce d d urin g ru n-in ,

Fo rtunate ly . a pro to type gearbox had beenrun at 1 /2 lo ad fo r about 5 0 hours. W hen thesegears w ere in peered , no signs o f d istress w erese en o n any o f the ge ar te eth . T he to ol.h pro filesw ere sm o oth , w ith su rfac e ro ug hn ess e stim a te dto be 20 j .J:inrrns, and t he c on ta ct p a tt er n i nd ic at ed1 0 0% f ac e c on ta ct. T h is g ea rb o x w a s r ea ss emb le da nd r un u nd e r 1 /2 lo a d u n til. i ts o il s ump - te mp era -lu re re ac he d e qu ilib riu m a t 2 0 0 F. F or th is a pp li-cat ion, th e am b ie nt te m pe ratu re w as in th e ra ng eo f 5 0 F to 1 25 F. T he c en te r d is ta nc e o f th e g ea rsw as 1 6 in ch es a nd th e p itch lin e v elo city w as 40 0 'fpm .. R eferring to A GM A 2 50 ..0 4, the recom -m e nd ed v isco sity fo r th ese co nd itio ns is IS O 1 .5 0 .or IS O 2 20 ..

Fig. 3 - Pressure-viscosity coefficient versus temperature for mineral gear lubricants ..(From AGMA 200I-B88. 1988.)

II

I - - - - - - I I: ~ - - - I

150 200 250BULKTEMPEAATURE, ( O F ) 30 0

U sin g th e e m pirical e qu atio n w e g et:

(3)7000V 40 "" : : : : :350 cSt

(400) 0.5

H ence, the em pirica l equation recom m ends av isc osity c lo se to I SO 3 20 .. It is a pp are nt th at th ev isco sity th at w as o rig in ally su pp lie d (IS O V G1(0 ) w as too low.

T he E HD film thickness w a s c alc ula te d w itha special computer program. 9 T he gear bulktemperature was assum ed to be 2 30 F (30 de-g re e s ho tt er than th e measured o il s ump t emp er a-ture) , The fo llow ing data fo r the ISO YO 1 . ' 0 0lu bricant w as o btain ed fro m F ig s. 2 and 3:

l J i .o= 6,6 cP(O .96 x 10 -6 Reyns)(J. = 1.0 2 x 1O-4in 21b

Fig ..4 shew s a plo t o f the film thicknessv ersus position on the pinion to oth . T he m ini-m um film th ick ne ss OCClUS low o n the pin i .on

35 0

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, t I S:gS ,.125t ' I : I ;, .J ,

Min lambda = . '073Probability of wear e>95% 5 .5

4 . 5.175

LPSTC

"/l.,1 3 . 5 '.. r

",,'"HPSTC

2.5

Pinion roll angle in degrees

fig. 4 - Film thick nes v ersus pinio n ro ll an gle fo r gear to oth geo metry o f cu ffed gearse t,

450

I

35'0

25 '0l . P STC HPSTC

15'0 1 - Max FIDash' Iernp = 439" ' " Scoring Probability e63%50 . I

Pinion r o n angle in degreesFig, 5 - Flash temperature versus pinion mil angle for gear tooth geometry of scuffed gearset,

2 0 GE"R T!ECHNOlOG'I'

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to m b n ear th e lowest p o in t o f s in gle to o th c on ta ct(LPSTC) where h . = 2 .1 m icro inches. ThemIDspecific film thickness, based on 2 0 uin rm ssu rf ac e ro ug hn es s f or b oth p ro file s, is A . ' " 0 . 073 .Fig. I. shows that th e g ears o pe rate in th e b ou nd -a ry lu br ic atio n re g im e . T h e p ro g ra m p re dic ts th atth e p ro ba bility o f w ea r is g re ate r th an 9 5 %.

T he c on ta ct te m pe ra tu re w as a ls o c alc ula te dw ith th e program. Th e s cu ff in g te mp era tu re fo rth e IS O B O 1 00 lu brican t w as c alcu late d w kh th ee qu atio n fo r n on -an ti- cu ff m in eral o ils:

Ts '" 146 + 59In(100); ; ; ; ; 418FF ig . 5 sh ow s a p lo t o f th e co m ac t te m pe ratu re

v ersu s p ositio n o n th e p in io n tooth, T he m axi-m um contact tem perature occurs high on thep in io n to oth n ea r th e h ig he st p oin t o f s in gle to othco ntact. (H PS TC ) w he re T c = 439F . The pro -gram pre dicts tha t the pro bability o f cuffing is63q(. This is considered to be a t high risk ofscu ffin g. T he re lativ ely h ig h te m pe ratu re p ea k:ne ar the tip o f the pinio n to oth w a caused by th eg eo m etr- y o f th e g ea rs. TIle d esig ne r se le cte d alo ng a dd en du m to oth for th e p in io n. L on g a dd en -dum pinions pe rfo rm w ell in speed reduce rs.w he re th ey in cre ase th e am o un t o f re ce ss ac tio na n d d ec re as e th e a m Oll ru o f a pp ro ac h a c tio n o f th e

Min lambda = .097Probability of wear = 94%

g ea r m e sh . S in ce re ce ss a ctio n :ism u ch s m oo th erthan appro ach actio n, lo ng adde ndum pinio nsgive s pe e d r ed uc ers s mo oth m e s hin g characteris-tics. W hen o pe ra te d as a speed increaser, ho w-e ve r, th e a pp ro ac h a nd re ce ss p ortio ns o fth e g ea rm esh rev erse , m ak ing a lo ng addend lim pinio nro ug h ru nn in g a nd v uln era ble to s cu ff in g.

T o e xp lo re th e p ossib ilitie s fo r reducing th escu ffin g risk . n ew g ear to oth g eo m etry w as p ro -p os ed w ith th e p in io n a nd g ea r a dd en da d es ig ne dto m in im iz e th e flash te m pe ra tu re rise . T he n ewgearset .analyzed w ith th e program, a ss um e d th elu brican t w as a m in era l o il with a n ti -s cu f f a ddi -t ives, with a v iscosity o f ISO 2 20 , and w ith thefo ll ow ing proper ti e s:

-6 R110 ::: 10 cP( 1.45 x 10 eyns)a = 1.09 x 10 -4 in 2 /lb

T s = 245 + 5 9In(2 20 ) '" 563"FFig, 6 sh ow s th at th e f ilm th ic kn es s in cre as es

to h _ = 2 .7 !lin , an d th e p ec ific film th ick ne ssmmi nc re a se s t o A . = 0 .0 9 7 .. Fig . I s ho w sth at th e g ea rs

s till o pe ra te in th e b ou nd ary lu bric atio n re gim e ,how ever, the probability o f w ear is reduced to94% , F ig . 7 sh ow s th at th e o ptim iz ed g ea r g eo m -e try re du ce d th e m a xim u m c on ta ct te m pe ra tu re toTc =3 0 2F . T he c om b in atio n o f re du ce d c on ta ct

. .275

//-/LPSTC HPSTC //. , . , . . . . - -.. . . , , -,-- - - - - .- - - - - - .. . . . - - -- - ~ ~ - - - - ~ - - -

.225

..125

. 075I: 20 40

Fig. 6 - Film thickness versus pinion roll angle for geartooth geometry optimized for maximum scuffing resistance.

30Pinion roll angle in degrees

6=r'3s: 3. . . . .("')4 "'I0::I

2

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350

HPSTC

250

LPSTC

20 I 30Pinion roll angie in degrees

temperature and the increased scuffing resistanceprovided by thehigher viscosity mineral oil withanti-scuff additives reduces (he scuffing prob-ability to < 5%.

Typical of many gear failures, this case his-tory shows that several factors contributed to thefailures:

.The lubricant viscosity was too low.-No anti-scuff additives were used.A gearbox designed as a speed reducer was

used as a speed increaser.The gear teeth were provided with a coat-

ing or plating to ease running-in,-The gears were not run-in properly under

reduced loads.Gear failures. as exemplified by the case

history, can be avoided if designers and opera-tors recognize that the lubricant is an importantcomponent of a gearbox, and appreciate that thetribology of gearing requires the considerationand control of many interrelated factors.

F ig . 7 - F lash te m pe ra ture v ersu s pin io n ro ll ang le fo r g ea r to oth g eo m etry o ptim iz ed fo r m axim u m sc uffin g resistance.

Max Flash Temp = 302

50 Scoring Probability =< 5%

7.DRAGO, R.J. Fundamentals ofGear Design . Butterworth,1988.8. AKIN, L. .&TOWNSEND, D. "Study of'Lubricant JetflowPhenomena in S pu r G ears ." N A SA T MX -7 IS7 2, O cL , 1974.

Acknowledgement: Reprinted by permission of Society of 9. SCORING+. Computer Program. GEARTECH Software,

References;1. .BLOK, H. "Les Temperatures de Surface dans les Cendi-l ions de Graissage sons Press ion Extreme," Second WorldPetroleum Congress, Paris, June, 1937.Z . AGMA 250.04. "AGMA Standard Specificat ion - Lubri-cation of Industrial Enclosed Gear Drives." Sept., 198l ,3. AGMA 421.06. "Practice for High Speed Helical andHerringbone Gear Units." Jan, 1969.4. DOWSON. D. "Elastohydrcdynumics," Paper No. 10,Proc. Inst , Mech.Engrs. , Vol. 182.PT3A. 1967.pp. 1.51-167..5. WELLAUER,EJ.&HOLLQWAY,G,A .."ApplicationofEHD Oil, Film Theory to Industrial Gear Drives." Trans.AMSEJ. Eng. Ind., Vol. 98, Series B. No.2, May. 1976. pp ,626-634.6. A_KAZAWA,M.,TEJ'IMA,T.& NARlTA, T. "FuIlSca!eTest of High Speed,High Powered Gear Unit - Helical Gearsof25,OOOPSat200m/sPLV." ASMEPaperNo.80-C2/DEl'-4.1980.

Tribologists and Lubrication Engineers. Inc., 1985-1989.

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