NASA Technical Memorandum 101430 

L Tribology: The Story of 

Lubrication and Wear 1 

Donald H. Buckley, William R. Jones, Jr., Harold E. Sliney 

Erwin V. Zaretsky, Dennis P. Townsend, 

and Stuart H. Loewenthal 

Lewis Research Center 

Cleveland, Ohio 

hemred for 

Se&nar F-107 at the 1985 International Trade Fair b Cleveland, Ohio, October 18, 1985 

k 

Ij ', 

(NASA-TH-101430) TRIBOLOGY: THE STORY OF 1309-24635 

LUBRICATION AND WEAR :NASA. Lewis Besearch --THBO-- 

Center) 145 p CSCL 20K N8 9 - 2 46 36 

Unclas 

G3/39 0204495 
CONTENTS 

Tribology 

Tribology Donald H. Buckley c 

............................................................................................. 3G,),, . 

(Published in “Advunced Materiak Technoiogy, ” NASA CP-2251, 1982.) gz#/,’=Pk7 

Lubrication 

Boundary Lubrication - Revisited 

William R. Jones, ........................................................................................ 82/V -?9f5-9 23(.j,., Jr. 

(Presented at Meeting of American socicry of Lubrication Engineers, 

Independence, Ohio, March 9, I982 (NASA 7u-82858).) 

Solid Lubricant Materials for High Temperatures-A Review 

Harold E. Sliney ............................................................................................... (Prepared for the Special Issue “Dry Beatings” of the Tribology International Jod, 

October 1982.) .. 

Mechanical Components 

Design and Lubrication of High-S&d Rolling-Element Bearings 

Erwin V. zaretsky .................................................................... f?~.~.<.?~~~ 85 

(Prcpamd for the Original Equipment Manufkturing Design Conference, Philodclphia, 

Pennsylvania, September 9-11, 1985 (NASA 7u-87107).) 

ssc/l/34LJL)7 101 

Lubrication and Cooling for High Speed Gears 

Dennis P. Townsend ......................................................................................... (Prrpand for the Original Equipment Manufucturing Design Conference, Philodclphia, 

Pennsylvania, September 9-11, 1985 (NASA liU-87096).) 

A Historical Perspective of Traction Drives and Related Technology 

Stuart H. Loewenthal ....................................................................................... ~93~~70 jJ3 119 -. ,c 

(Published in Adwnccd Powr Tmmission Technology, NASA CP-2210, 1981.) 

i 
Tribology 

I 
TRIBOLOGY* 

b 

c 

Donald H. Buckley 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

Tribology is the study of the adhesion, friction, wear, and lubricated 

behavior of materials in solid-state contact. The function of tribological 

research Is to bring about a reduction in the adhesion, friction, and wear of 

mechanical components to prevent their failure and provide long, reliable component life through the judlcious selection of materials, operating parameters, 

and lubricants. 

Mechanical systems such as bearings, gears, and seals are examples of 

components involving tribology. Wherever, however, two or more solid surfaces 

are in contact with relative motion between the surfaces, tribology is involved. 

Such mundane activities as a man's morning shave involve both friction and 

corrosive wear, and considerable tribological research has gone into increasing 

blade life and reducing friction and shaving discomfort. More complex tribological systems are gyro bearings and instrumentation gears requiring attention to many elements. 

and wear properties of materials and some of the factors influencing these 

properties. The forms of lubrication and types of lubricants will also be 

discussed. 

The objective of the present paper is to review the adheslon, friction, 

ADHESION AND FRICTION 

In a conventional atmospheric environment, the oxygen present In the air 

interacts with freshly generated metal and alloy surfaces to produce surface 

films; namely, oxides. These oxides play a very strong role in the adhesion, 

friction,and wear behavior of metals and alloys. In the absence of these 

oxides, very strong adhesion, high friction coefficients, and ultimately cold 

welding of materials from one surface to another are observed. If, for 

example, two normal metal surfaces are placed Inside a vacuum environment with 

a system capable of achieving pressures to 10-10 torr, and the surfaces are 

very carefully cleaned with argon ion bombardment and then brought into touch 

contact, adhesion of .one surface to the other will immediately occur. Attempts 

at tangential motion will cause a growth in the area adhered at the interface, 

that is, in the real contact area with an ultimate complete seizure of the 

surfaces one to another. With this occurrence, the surfaces are generally 

severely disrupted (refs. 1 and 2). 

When adhered surfaces are separated, the adhesion at the interface between 

the two dissimilar solid surfaces is sufficiently strong to prevent fracture 

at the Interface, but fracture does generally occur in the cohesively weaker 

*Published In "Advanced Materials Technology," NASA CP-2251, 1982. 

PRECEDiNG PAGE BLANK NOT FILMED 3 
of the two materials in contact (ref. 3). An example of such behavior is shown 

in the photomicrograph of figure 1. This figure represents the results of an 

experiment conducted In a vacuum chamber where two solid surfaces were brought 

into contact; the surfaces had been.cleaned in vacuum. Adhesion occurred at 

the interface, and when separation of the solid surface was attempted, fracture 

occurred in one of the two materials, leaving material transferred to the 

opposite surface. 

The actual area of real contact at the interface can be seen in figure 1 

by a careful examination of the interface region. There appear to be voids in 

the interface region where complete and Intlmate solid-state contact across the 

interface did not occur. However, a great portion of the interfacial area does 

reflect solld-to-solid contact in adhesive bonding. This adhesive bond was 

generated as a result of attempting to slide one surface over the other. 

resulted In growth In the adhered junctions at the interface, leaving only a 

small are wherein no intimate contact of the two solid surfaces occurred. On 

separation, the adhesive bond at the Interface remained intact and fracture 

occurred in one of the solids, as Indicated in figure 1 by the rough piece 

which remained on the solid surface. 

This 

This type of adhesion and transfer behavior is observed for all clean 

metal surfaces and alloys in solid-state contact. That is, when the environmentally contributed surface films, namely the oxldes, are removed, such adhesion Is observed with strong bonding and friction coefflclents measured in 

excess of 100 under such circumstances. The surfaces of solids such as metals 

and alloys are so sensitive to the contact region microenvironment that the 

admission of very small adsorbate concentrations from the environment to the 

solid surface can markedly reduce adhesion and friction. For example, fractions of a monolayer on the solid surface will produce a marked reduction In 

the adhesion and corresponding static friction coefficients for metal in 

contact (ref. 4). 

influenced by the presence of the environment in their adheslon, frictlon, and 

wear behavior. The presence of adsorbates on ceramic materials such as aluminum oxide has a pronounced influence on the friction coefficients measured for 

aluminum oxide (ref. 5). 

In addition to metals and alloys, nonmentallic materials are also markedly 

Polymeric materials are also affected by the presence of environmental 

constituents on their surfaces. For example, nylon serves as a good solid 

self-lubrlcanting material In certain mechanical applications. Nylon, however, 

depends upon the presence of adsorbed moisture for its effective lubrication; 

that is, for Its low friction and wear properties. In the absence of moisture, 

nylon will not lubricate effectively and it becomes a poor tribological material (ref. 5). Carbon materials are heavily used in mechanical devices such 

as dynamic seals and are extremely sensitive to environment and environmental 

constituents. 

The flying of aircraft at high altitudes results in excessive wear ln 

carbon generator brush materials; this was established during the second World 

War. Careful analysis of the aircraft surfaces revealed that the excessive 

wear of carbon materials, carbon bodles at high altitudes, was due to a reduction in the ambient pressure, and more partlcularly with the reduction of 

moisture in the environment. Carefully controlled experlments In the laboratory subsequently demonstrated that carbon lubricates effectively in the 

4 
C 

presence of moisture, exhibiting low ftictlon, low wear, and little tendency 

to adhere. In the absence of moisture, however, carbon exhibits extremely 

heavy wear and becomes a very p o o r f r l c t i o n a n d wear material. 

simply reducing the ambient pressure from 760 torr of alr to an amblent pressure of approximately 1 torr, a 1,000-fold change In wear properties occurs. 

Thus, moisture is a lubrlcant and is needed on these surfaces and at the interface between two carbon bodies in relative contact and in motion, or between 

the carbon body and some other material in solid-state contact (ref. 5). 

In fact, by 

If one considers the environment not only as an ordinary air containing 

principally oxygen and nitrogen w i t h s o m e w a t e r vapor, but also conslders 

vapors of hydrocarbons as constituents of the environment, then the particular 

hydrocarbon molecular structure that may be present in the environment can have 

a very pronounced infl'uence upon the adhesion, friction, and wear behavior of 

materials in contact. For example, careful cleaning of iron surfaces in a 

vacuum environment will result in the generation of extremely energetic surfaces that wlll adhere one to another when brought into contact with cold welding occurring readily. If, however, a small amount of hydrocarbon gas is 

admitted to the vacuum chamber and allowed to absorb on the clean iron surface, 

a structure wlll develop which will provide that s u r f a c e w i t h a protective film. 

Further, clean iron surfaces wlll chemisorb nearly all hydrocarbons 

(ref. 6). The hydrocarbon film will reduce adhesion, frlction, and wear 

because the surface energy has been reduced by the hydrocarbon molecules on the 

surface. The energy on the clean iron surface available for bonding across the 

interface to another solld surface has been taken up in the interaction of the 

clean Iron s u r f a c e w i t h the l u b r i c a t i n g m o l e c u l e s absorbing on It. The part i c u l a r m o l e c u l a r structure, however, of the adsorbing hydrocarbon wlll also 

affect the tribological behavior. That is, a slight modification in the molecule will produce sensittvities in adhesion, friction, and wear. These slight 

d l f f e r e n c e s i n d i c a t e e x t r e m e ' s e n s i t i v i t y in the tribological behavtor of materials to environment and envlronmental constituents. This effect can be demonstrated by the adsorption af a sfmple hydrocarbon such as ethylene oxide onto 

an iron surface and exposure of that same surface to a different simple hydrocarbon with a sl'ightly modified molecular structure, something such as ethylene 

chloride ( o r a s it is commonly called, vinyl chloride). 

If clean iron single crystal surfaces of the same orientation are exposed 

to equivalent concentrations of these two simple hydrocarbons, namely ethylene 

oxide and vinyl chloride, entirely different surface structures result. The 

differences can be seen ln the LEED (Low Energy Electron Diffraction) patterns 

presented in figure 2. 

LEED Is a d e v i c e w h i c h permits examlnatIon of the structural arrangement 

of atoms In the outermost atomic layer of the solid surface. Thus, in 

figure 2, we see the molecular arrangement ln the diffraction pattern for the 

adsorbed ethylene oxide and vinyl chloride on the Iron surface in the two patterns. Equivalent concentrations of each specie were provided. Thus, everything is constant except the particular molecular structure. The ethylene 

oxide exhibits the basic ethylene structure with oxygen present in the molecule. Vinyl choloride exhibits essentially the same structure as the ethylene, 

but chlorine Is substituted for oxygen. This subtle difference in the structures, however, causes marked differences in surface coverage (see LEED 

patterns, fig. 2). 

5 
With the ethylene oxide, the six diffraction spots in a hexagonal array 

indicate that the ethylene oxide molecule completely masked or covered the iron 

surface. No diffraction spots are seen for the Iron in the diffraction pattern of figure 2. A close packing of the molecules of ethylene oxide on the 

iron surface provides a very effective and continuous surface film. 

In contrast to ethylene oxide, however, the vinyl chloride structure is 

Much more open with less-than-complete surface coverage and bonding of vinyl 

chloride to the surface. The four bright diffraction spots with vinyl chloride 

adsorption, seen in a rectangular array in figure 2, are associated with the 

iron. Thus, vinyl chloride does not provide complete coverage, and nascent 

iron is still exposed at the surface. 

As one might anticipate, differences In adhesion and friction behavior are 

observed with these two films present. With the ethylene oxide, the adhesive 

forces are appreciably reduced between two clean iron surfaces In contact. 

Further, the friction forces are less for the ethylene oxide on the iron 

surface than are observed with the vinyl chlorlde present. Thus, slight differences in the molecular structure of hydrocarbons present in the microenvironment of solid surfaces in contact can influcence'the tribological 

behavior of those surfaces. 

WEAR 

Various recognized mechanisms cause surfaces to wear. The more common 

types are adhesion, abrasion, corrosion, erosion, cavitation, fretting, and 

fatigue. Some of these mechanisms will be briefly discussed. 

Adhesive Wear 

Adhesion has already been discussed with reference to figure 1. The 

transfer resulting from the interfacial adhesion Is adheslve wear. Material 

has been lost from one surface and transferred to another. Adheslve wear can 

occur for a wide variety of materials brought Into contact. Abrasive wear, 

however, is limited to those situations where a very hard material contacts a 

softer matertal, or where hard particles are sandwiched between two softer 

surfaces (e.g.', particles of sand in a bearing). 

surface is cut or mlcromachined by a harder surface or particle. 

Abrasion occurs when a softer 

Abrasive Wear 

One might intuitively anticipate that the resistance of a material to 

abrasive wear is strongly a function of the hardness of the surface being 

abraded. The harder the surface, the greater should be the resistance to 

abrasion; this has been experimentally demonstrated (ref. 7). 

In figure 3 resistance to wear Is plotted as a function of hardness for 

the surface of various metals. The data of figure 3 indicate a direct relation 

between the hardness of the metal-being abraded and Its abraslve wear 

resistance. 

6 
The abrasion of solid surfaces Involves wear to the abrasive as well as 

wear to the surface being abraded. For example, wlth such relatively hard 

abraslve materials as single crystal aluminum oxide (sapphire) and titanium 

dloxide (rutile), the resistance to wear 1s very much a function of their 

orientation. With certain atomic planes contacting steel, wear reslstance of 

the abrasive substance is greater than for other orientations. This resistance 

Is demonstrated in the data of figure 4 for titanium dioxide. 

A marked variation in the wear rate of titanium dioxide (fig. 4) occurs 

with changes in orientation. Between the minimum and maximum it varies by a 

factor of seven times. Thus, abrasion can result in wear to the abrasive as 

well as to the surface being abraded; the latter can be minjmlzed by giving 

consideration to the physical and mechanical properties of the abrasive 

materlal. 

Corrosive Wear 

The surfaces of solids play an extremely important role in corroslve wear. 

In corrosive wear, material is lost fr0m.a solid as a direct result of chemical 

interactions of the solid surface with the environment. The active environmental constituent can be the lubricant, an additive, or a component of the 

surrounding atmosphere. The relative motion between solid surfaces in contact 

aggravates surface attrition by continuously exposing fresh surface for 

reaction. 

Materials which are very effective lubricants under certaln condlttons can 

become extremely reactive under another set of conditions. The l u b r i c a t i o n o f 

alloys wlth halogen-contalning 1ubri.cants is a good example. 

In flgure 5 for a cobalt alloy lubricated by a chlorlnated fluorocarbon, 

wear at temperatures to 300 "C is extremely low. The values in figure 5 are 

100 times less than those obtained for the unlubricated surfaces. Above 

300 "C, however, the rate of wear begins to increase markedly. This increase 

is due to excessive chemical reactivity of the chlorine of the chlorinated 

fluorocarbon wlth the cobalt surface. 

Examination of the cobalt alloy surface after sliding revealed copious 

quantities of cobalt chlorlde. This particular compound is an extremely good 

solid-film lubricant and accounts for the low wear to 300 "C. Above that temperature cobalt chloride contlnues to form, but in such large quantities that 

the cobalt alloys are consumed as a result of excesslve surface reactlvlty. 

Thus, effective lubrication, with solid films of the type described here, is a 

matter of controlled corroslon. A reaction product should form to reduce 

frlction and wear, as In figure 5, but that quantlty should be limlted. 

The data of figure 5 also Indicate that no correlation between friction 

and wear can be drawn from information about one or the other. Corrosive wear 

is an excellent example of this concept. Wear may increase due to the excessive reactivlty, but friction may decrease because of the low shear strength 

of the reaction product formed. 

Corrosive wear can be brought about by increasing temperature, as 

lndlcated In figure 5. Slmllar effects can be produced by both Increased 

loading and increased rubblng speeds. 

7 
A wide variety of material properties affects wear behavior. As already 

indicated with reference to figure 4, the crystallographic orientation of 

materials affects wear. This property holds true not only for the wear of 

nonmetals, but for metals as well (ref. 8). 

Crystal structure is another property of materials which influcences wear. 

Transformation in a metal from one crystal structure to another can result in 

notable changes in wear. This effect is indicated for t i n I n figure 6. The 

wear track width is plotted as a function of temperature, and is realtively 

constant until the temperature for the transformation of tin from a diamond 

structure (gray tin) to that of the tetragonal (white tin) is approached. At 

that point wear begins to Increase. The tetragonal tin structure has greater 

ductility than the diamond form. 

A further manifestation of the effect of crystal structure is observed 

when layer lamellar solids are In rubbing contact with metals. With these 

solids shear readily occurs along basal planes, and transfer to the metal surface is readily observed; this can be seen In table I. In the table pyrolytic 

boron nitride transfers to all metals except gold and silver; poor adhesion 

accounts for this failure. 

In practical englneerlng~applications, metals are not used in their 

elemental form, but rather as alloys. The alloying elements can have varying 

effects on wear. For example, In figure 7, adding 10-atomlc-percent aluminum 

to copper does not affect its rate of wear. The addition of 10-atomic-percent 

of alloying element such as silicon, tin, or indium to copper does, however, 

appreciably reduce wear (fig. 7). 

The wear behavior differences for copper alloyed with. various elements are 

maintained even with variations in the concentration of the lubricant additive, 

as is Indicated in figure 8 for the alloys copper-70-atomlc percent aluminum 

and copper-10-atomic-percent indium. At all concentrations of stearic acid, 

wear is greater with aluminum alloyed with copper than it is for indium alloyed 

with copper. 

LUBRICATION 

Liquids 

The purpose of lubrication Is to separate surfaces in relative motion by 

a material which has a low resistance to shear so that the surfaces do not 

sustain major damage. This low-resistance material can be any of a variety of 

different species (e.g., adsorbed gases, chemical reactions fllms, liquids, 

solid lubricants), some of which have already been discussed. 

Depending on the type of intervening film and Its thickness, a number of 

lubrication regimes can be identified. A classical way of depicting some of 

these regimes is by use of the well-known Strlbeck curve (ftg. 9). Stribeck 

(ref. 9) performed comprehenslve experiments on journal bearjngs around 1900. 

He measured the coefficient of friction as a function of load, speed, and temperature. He had difficulty, however, condensing this data into usable form. 

Some years later, Hersey (ref. 10) performed similar experiments and devised a 

plotting format based on a dimensionless parameter. The Stribeck curve, or 

more appropriately, the Stribeck-Hersey curve, takes the form of the 

8 
c 

coefficient of friction as a function of the viscosity of the liquid (Z), 

velocity (N), and load (P) parameter, ZN/P. 

At high values of ZN/P which occur at high speeds, low loads, and at high 

viscosities, the surfaces are completely separated by a thick (>0.25 pm) 

(>lO-sin.)) lubricant film. 

where friction is determined by the rheology of the lubricant. For nonconformal concentrated contacts where loads are high enough to cause elastic 

deformation of the surfaces and pressure-viscosity effects on the lubricant, 

another fluid film regime, elastohydrohynamic lubrication (EHL), can be 

identified. In this regime film thickness (h) may range from 0.025 to 2.5 pm 

(10-6 to 10-4 in.). 

As film thickness becomes progressively thinner, surface interactions 

start taking place. This regime of increasing friction, which combines 

asperity interactions and fluid film effects, is referred to as the mixedlubri cation regime. 

This area is that of hydrodynamic lubrication 

Finally, at low values of the ZN/P parameter, one enters the realm of 

boundary lubrication. Thls regime is characterized by the following: 

1. This regime is highly complex, involving metallurgy, surface topography, physical and chemical adsorption, corrosion, catalysis, and reaction 

kinetics. 

tive surface films to m i n i m i z e w e a r and surface damage. 

2. The most important aspect of.this regime is the formation of protec3. The formation of these films Is governed by the chemistry of the filmforming agent, as well as the.surface of the solid and other environmental 

factors. 

4. The effectiveness of these films in minimizing wear is determined by 

their physical properties, which include shear strength, thickness, surface 

adhesion, film cohesion, melting point or decomposition temperature, and 

solublllty. 

Besides the Strlbeck-Hersey curve (fig. 9) already described, an idealized 

plot of wear rate as a function of relative load can also delineate the various 

lubrication regimes and some wear transitions (fig. 10, ref. 11). 

Region OA of figure 10 encompasses the regimes of hydrodynamic and EHL, 

the latter as point A is approached. Since no surface interactions occur in 

this region except for startup of shutdown, little gr no wear occurs. (This 

excludes rolling-element fatigue, which can occur wlthout surface interactions.) Region AX is the mixed-lubrication r e g i m e w h e r e surface interactions 

begin to occur at A and become more prevalent as point X is approached. 

is low because fluid film effects still exist. 

Wear 

Next there is region XY in figure 11, which is the region of boundary 

lubrication. The degree of metal-to-metal contact and the wear rate increase 

as the load increases. Wear is mild and tends to be corrosive to the left of 

B and adhesive to the right of B. The locatlon of B Is quite varlable and 

depends on the corrosivity of the lubricant formulation. For a noncorrosive 

lubricant, adhesive wear can occur at X. On the other hand, a corrosive 

9 
additive can extend the boundary regime to Z1 before boundary film failure 

occurs. Region YZ is the regime of severe wear where severe adhesion and 

scoring occur. Machinery cannot operate successfully in this region, and, 

therefore, the location of this transition point is quite Important. At point 

Z total surface fallure occurs, followed by seizure. 

In the boundary lubrication regime many properties of the liquid lubricant 

become important. These include shear strength, film thickness, melting point, 

and chemical reactivity with the surface. Operating variables which will 

affect lubricant film-performance include load, speed, temperature, a.nd atmosphere, as already discussed. Additives present in the lubricant to serve 

specific functions will also affect behavior. These additives include antiwear, antifoam, antloxidants, viscosity improvers, and others. A good review 

of boundary lubrication can be found in reference 12. 

Sol ids 

At temperatures below which liquid lubricants become solid, and above 

which they either thermally or oxidatively decompose', solids are used. 

solids include Inorganic compounds, polymers, and low-shear-strength metals. 

A review of the subject can be found In reference 13. 

The 

Of solid lubrlcants, those most widely used and studied are the layerlattice inorganic compounds. These materials have a hexagonally layered 

crystal structure. Thelr shear properties are anisotropic with preferred 

planes for easy shear parallel to the basal planes of the crystallites. In 

some of the compounds such as molybdenum disulfide (MoSz), a low shear 

strength is Intrinsic to the pure material, which in others, notably graphite, 

the presence of absorbed gases or intercalated "impurltles" between the basal 

planes appears to be necessary to develop desirable friction characteristics. 

The most common representatives of this class of lubricants are graphite and 

the dichalcogenldes, notably MoS2 and WS2. 

The maximum useful temperatures for solid lubricants depend strongly upon 

the composition of the ambient atmosphere, the required life at temperature, 

factors such as oxygen availability at the lubricated surface (is the coating 

openly exposed to the atmosphere, or shielded within conformlng bearing surfaces?), air flow rates, lubricant particle size, and the Influence of 

adjuvants and binders. 

Dichalcogenides 4 

The maximum temperature for lubrication with MoS2 In an air atmosphere 

is limited by oxidation to about 400 OC under favorable conditions. Some .I 

oxidation kinetics data for loosely compacted MoS2 powders of lp average 

particle size are given in figure ll(a) from reference 13. At a modest airflow 

rate over the compact, 50 percent of the MoS2 was oxidized to molybdlc oxide 

(Moog) in 1 hr at 400 "C. At a six-times-higher airflow rate, the temperature for an oxidation half-life of 1 hr was reduced to 300 "C. Figure l(b) 

compares the oxidation kinetics of MoS2 and WS2 at the lower alrflow rate. 

The curves for MoS2 and WS2 intersect, with MoS2 oxidizlng more rapldly 

above about 340 "C. 

10 
Friction experiments were conducted with a pin-on-disk apparatus using a 

hemlsphertcally tipped pin In sliding contact with the flat surface of a 

rotating disk. A comparison of the oxidation data of figure 11 and the 

friction data of figure 12(b) (ref. 13) shows that the loss of lubricating 

ability of MoS2 and WS2 in air coincides with the temperatures at which 

rapid conversion to the oxides occurs. 

Figure 12(a) also shows that both compounds lubricate to much higher temperatures in a nonreactive argon atmosphere. In an inert gas or vacuum, the 

maximum useful temperature is a function of the thermal dissociation rates, 

rather than the oxidation rates of the lubricants. Thermal dissociation rates 

and the friction coefficients of molybdenum and tungsten disulfides, dlselenides, and ditellurides in vacuum have been systematically studied (ref. 14). 

The major results, summarized In table 11, indicate that the disulfides are the 

most stable, the diselenides are intermediate, and the ditellurides are the 

least stable. However, thin, burnished films of the dlselenides with their 

higher d e n s i t i e s e v a p o r a t e m o r e slowly than the disulfides. Apparently, for 

the very thin, burnished films, the evaporation rates were the controlling 

factor In determining the maximum temperature for effective lubrication. The 

limiting temperatures for a significant wear life of these coatings ranged from 

600 to 700 OC. 

Vacuum-deposited coatings are Increasingly being used in tribological 

applications; these fall into two rnaln composition categories: soft-lubrlcattng coatings, and very hard, wear-reslstant coatings. The methods of appllcation are also in two principal categories: sputtering and ion-platlng. These 

techniques have been rapidly adopted by industry especially for aerospace 

applications. A very large variety of these vacuum-deposited coatings is 

becoming available for lubrication application. 

The most common vacuum-depos1,ted tribological coatings are the sputtered 

Compounds such a s H o S 2 a r e usually applied by sputtering, because with 

dichalogenides, especially MoS2 and ion-plated soft metals such as gold. 

These coatings are often very thin, on the order of 2000 to 5000 A in thickness. 

proper procedures, pure, essentially stolchlometrlc, compounds can be deposited. 

In contrast, ion-plating tends to dissocate chemical compounds. However, ionplating is an appropriate technlque for the deposltlon of elemental metals 

because (1) dissociation Is obviously not a problem; (2) hlgh lon-impact 

energies can be used to enhance adhesion; (3) excellent throwing power is 

achieved when coating parts with complex shapes; and (4) rapid deposition rates 

tan be achieved. 

Sputtered hard coats are used primarily for wear control. The oxidation 

temperatures and hardness of some important carbides and nltrides are presented 

in table 111. Coatings of .all of the compounds listed are hard enough to be 

expected to have good wear resistance, assumlng adequate bond to the substrate 

can be achieved. However, a considerable variation In oxidation resistance 

exists. Chromium carbide. boron carbide, silicon nitride, and sillcon carbide 

are oxidatively stable to at least 1000 "C, while tungsten and titanium carbldes oxidize durlng long-duratlon exposure in air at temperatures above about 

540 OC. Tungsten carbide tends to oxidize more rapidly than titanium carbide 

because its oxides are volatile at high temperature, and their sublimation 

tends to accelerate the oxidation. 

11 
Titanium nitride Is another promising hard-coat material, but it too will 

convert to the oxide above 550 OC. However, some TiC- and TiN-sputtered 

coatings have shown surprisingly good resistance to oxide conversion at hjgher 

temperatures than those listed In table 111. Oxidation occurs, but the rate 

is very low, probably because of high coating density and the passivating 

nature of the initially formed oxide films, which protect the coating against 

catastrophic oxidation. 

CONCLUDING REMARKS 

The adhesion, friction, and wear behavior of materials in solid-state 

contact are strongly dependent upon material properties and environmental 

factors. Adsorbed surface films and oxides on metals markedly influence 

trlbological behavior. 

Relatively subtle differences in the molecular structure of hydrocarbons 

can profoundly influence adhesion and friction. With liquid lubrication, 

distinct regimes of lubrication are identified. Both the physical properties 

of the lubricant and operating variables affect lubricant performance. 

With solids as high-temperature lubricants, both thermal and oxidative 

stability are important, as well as the lubricating properties of the solids 

over a broad temperature range. Techniques such as ion-plating and sputtering' 

are being increasingly used for the application of solid film lubricants. 

REFERENCES 

1. Buckley, D.H.; and Johnson, R.L.: Friction and Wear of Hexagonal Metals 

and Alloys as Related to Crystal Structure and Lattice Parameters in 

Vacuum. ASLE Trans., vol. 9, no. 2, 1966, pp. 121-135. 

2. Tabor, 0.: Junction Growth In Metallic Friction: The Role of Combined 

Stresses and Surface Contamination. Proc. Roy. SOC. (London), A, 

vol . 251, 1959. 

3. Buckley, D.H.: Definition and Effect of Chemical Properties of Surfaces 

In Friction, Wear and Lubrication. NASA TM-73806, 1978. 

4. Wheeler, D.R.: Effect of Adsorbed Chlorine and Oxygen on the Shear 

Strength of Iron and Copper Junctions. J. Appl. Phys., vol. 47, no. 3, 

Mar. 1976, pp. 1123-1130. 

5. Buckley, D.H.: Friction, Wear and' Lubricating in Vacuum. NASA SP-277, 

1971. 

6. Buckley, D.H.: Adsorption of Ethylene Oxide and Vinyl Chloride on an 

Iron (011) Surface and Effect of these films on Adhesion. NASA 

TN 0-5999, 1970. 

7. Khruschov, M.M.: Resistance of Metals to Wear by Abrasion, as Related to 

Hardness. Proceedings of the Conference on Lubrication and Wear. 

Institution of Mechanical Engineers, London, 1957, pp. 655-659. 

12 
8. 

9. 

10. 

11. 

12. 

13. 

14. 

Buckley, D.H.; and Johnson, R.L.: Friction and Wear of Hexagonal Metals 

and Alloys as Related to Crystal Structure and Lattice Parameters in 

Vacuum. ASLE Trans., vol. 9, 1966, pp. 121-135. 

Stribeck, R.: Characteristics of Plain and Roller Bearings. Z.V.D.I., 

vol. 46, (1902). 

Hersey, M.D.: 

J. Wash. Acad. Sci., vol. 4, 1914, pp. 542-52. 

The Laws. of Lubrication of Horizontal Journal Bearings. 

Beerbower, A.: Boundary Lubrication. Scientific and Technical Applications Forecasts, Department of the Army, DAHC-19-69-C-0033, 1972. 

Jones, W.R., Jr.: Boundary Lubrication Revisited. NASA TM-82858, 1982. 

Sliney, H.E.: Solid Lubricant Materials for High Temperature - A Revlew. 

Dry Bearings, Tribology International Journal, 1982. 

Brainard, W.A..: The Thermal Stability and Friction of the Disulfides and 

Ditellurides of Molybdenum and Tungsten In Vacuum. 

torr). NASA TN 0-5141, 1969. 

to 

13 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

CS-322l8 

Figure 1. - Severe surface welding resulting from unlubrioted sliding. 

(2% AI-Ni alloy from 10-9 mm Hg vacuum experiment) 

14 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

0 /\ 

CH2 - CH2 

CHZ = CHCl 

Figure 2 - Leed patterns obtained with huo polymer forming 

hydrocarbons on iron (011) surface. 100 L exp. 

15 
40 

32 

a 

Pb I I I 

0 80 160 240 320 

Hardness, kglmm' 

Figure 3. - Resistance to wear as function of hardness 

(ref. 7). 

Mo 

8x10'4 

r 

5 :I 4 

0 n 

0 

Polar angle, de9 

Figure 4. - Rate of wear of rutile single-crystal 

sphere on great circle in plane of a- and caxes. The c-axis is normal to plane of sliding 

at 8 and 1800. Sliding direction is in plane of 

great circle. 

16 
.4 

.2 

0 

I I I I I 

100 200 300 400 m 

Temperature, OC 

Figure 5. - Friction and wear of cobalt alloy sliding on itself at various temperatures and lubricated with 

chlorinated fluoroarbpn. 

rTin crystal 

I,,’ transformation 

I I tr I I I 

-150 -100 -50 0 50 1M1 150 

Temperature. OC 

Figure 6. - Track width on tin single-crystal surface as function 

of temperature SI ing velocity. 0.7 mmlmin: load, 10 9: 

pressure, 10-8’ Nlm !t , rider. iron (110): single pass.. 

17 
10-2 

L 

c E N31 

E 

a 

u 

104 L 

.- 

VI & 

0 

E 

VI 

L 

-t 

m 

1 

a 

u 

n 

E 

rc 

- 

-r m 

1 

a 

1 

Figure 7. - Rider wear for various copper alloys sliding 

on themselves in hexadecane containing 0.1 vol 70 

stearic acid. Load, 2M8; sliding velocity. 300 cml 

min; temperature, b C. 

'"a OCu-lOatmAl 

0 Cu-10 at % In 

- l*k- - 

0 .05 . 10 . 15 e20 .25 .30 

16 

Stearic acid in hexadecane, percent 

Figure 8. - Rider wear rate for two copper alloys sliding on themselves 

with various concentrations of stearic acid in hexadecdne as lubricant. Load. 500 g; sliding velocity, 300 cmlmin; temperature, 25 OC. 

18 
hzgzid- I+ 

I! I h > loa5 in. 

Transition 

3 

- 0 

e 

/ 2' 

00' 

I I i 

(viscosityl (velocity) 

load 'P 

Figure 9. - Coefficient of friction as function of viscosityvelocity-load parameter (Stribeck-Hersey curvet (ref. 9). 

Seizure 

Severe 

Transition / 2' 

00' 

/MO 

Boundary 7 

i I 

0 /' AX B Y 2 

'Hydrodynamic Relative load 

Figure 10. - Wear rate as function of relative load depicting various regimes of lubrication (ref. 11). 

19 
Airflow 

rate, 

cm3/min 

- 333 

am 

- 1 Week 

-\. 

104 

Id 

102 

10 

1 

- 1 Week 

I 

600 

Temperature, OC 

(a) Oxidation characteristics of MoS at two airflow rates. 

(M Comparative oxidation of WS2 an? MoS2 at airflw rate 

of 113 literlmin. 

Figure 11. - Oxidation kinetics of MoS and WSp 

Average particle size, 1 vm ; compacfdensily. 

50 percent. 

ows; 

.2 - 0 

I I I 

" 

Temperature, OC 

Figure 12 - Friction characteristics of MoS2 and WS2 in argon 

(a) Argon atmosphere. (b) Air atmosphere. 

and in air. 

20 
Lubrication 

Liquids 

Solids 

21 
BOUNDARY LUBRICATION - REVISITED* 

William R. Jones, Jr. 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

A review of the various lubrication regimes, with particular emphasis on 

boundary lubrication, is presented. The types of wear debris and extent of 

surface damage is illustrated for each regime. The role of boundary surface 

films along with their modes of formation and important physical properties are 

discussed. In addition, the effects of various operating parameters on friction and wear In the boundary lubrication regime are considered. 

INTRODUCTION 

The purpose of lubrication is to separate surfaces In relative motion by a 

material which has a low resistance to shear so that the surfaces do not sustain major damage. This low resistance material can be a variety of different 

species (e.g., adsorbed gases, chemical reaction films, liquids, solid lubricants, etc .) . 

Depending on the type of intervening film and its thickness, a number of 

lubrication regimes can be identified. A classical way of dsplcting some of 

these regimes Is by use of the well known Stribeck curve (fig. 1). Stribeck 

(ref. 1) performed comprehensive experiments on journal bearings around 1900. 

He measured the coefftcient of friction as a function of load, speed, and temperature. However, it was difficult to,condense this data into usable form. 

Some years later, Hersey (ref. 2) performed similar experiments and devised a 

plotting format based on a dimensionless parameter. The Stribeck curve, or 

more appropriately, the Stribeck-Hersey curve takes the form of the coefficient 

of friction as a function of the viscos4ty (Z), velocity (N), and load (P) 

parameter, ZN/P. 

viscosities, the surfaces are completely separated by a thick (>0.25 pm) 

(>10-5 In.) lubricant film. This Is the area of hydrodynamic lubrication where 

friction is determined by the rheology of the lubricant. For nonconformal concentrated contacts where loads are high enough to cause elastic deformation of 

the surfaces and pressure-viscosity effects on the lubricant, another fluid 

film regime, elastohydrodynamic lubrication (EHD) can be Identified. On this 

regime film thicknesses (h) may range from (0.025 to 2-5 pm) (10-6 to 10-4 in.). 

At high values of ZN/P which occur at high speeds, low loads, and at high 

As film thickness becomes 

takIng place. This regime of 

tion of asperity interactions 

mixed lubrication regime. 

*Presented at Meeting of Amer 

progressively thinner, surface interactions start 

increasing friction In which there is a combinaand fluid film effects is referred to as the 

can Society of Lubrication Engineers, Independence, Ohio, March 9 , -1 982 (NASA TM-82858). 

PRECEDING PAGE BLANK NOT FILMED 

23 
Finally, at low values of the ZN/P parameter, one enters the realm of 

boundary lubrication, the primary subject of this paper. This regime is characterized by the following (ref. 3): 

1. It is a highly complex regime involving metallurgy, surface topography, 

physical and chemical adsorption, corrosion, catalysis, and reaction kinetics. 

2. The most important aspect of this regime is the formation of protective 

surface films to minimize wear and surface damage. 

3. The formation of these films is governed by the chemistry of both the 

film former as well as the surface and other environmental factors. 

4. The effectiveness of these films in minimizing wear Is determined by 

their physical properties which include: shear strength, thickness, surface 

adhesion, film cohesion, melting point or decomposition temperature, and 

solubility. 

At is obvious that a conclse definition of boundary lubrication is not 

possible. For the purposes of this vaper, the following general definition 

will be used. Boundary lubrication 1s lubrication by a llquid under conditions 

where there is appreciable solid-solid interactions. Friction and wear are 

determined predominantly by interactions between the solid surfaces and between 

the surfaces and the liquid. The viscous properties of the liquid play little 

or no part in this process. 

The purpose of this paper Is to summarize the present knowledge about 

boundary lubrication. 

the literature. A number of adequate surveys already exist (refs. 3 to 8). 

There is no intent to provide an exhaustive survey of 

Lubrication Regimes 

Besides the Stribeck-Hersey curve (fig. 1) descrlbed In the introduction, 

an idealized plot of wear rate as a function of relative load can also delineate the various lubrication regimes as well as some wear transitions (fig. 2) 

(ref. 8). 

Region OA. - This region encompasses the regimes of hydrodynamic and 

elastohydrodynamic lubrication (EHD), the latter as point A is approached. 

Since there are no surface interactions in this region except for startup or 

shutdown, little or no wear occurs. This is excluding rolling element fatigue 

which can occur without surface Interactions. 

Region AX. - Region AX is the mixeu lubrication regime where surface 

interactions begin to occur at A and become more prevalent as point X is 

approached. Wear is low because there are still fluid film effects. 

Region XY. - This is the region of boundary lubrication. The degree of 

metal to metal contact and the wear rate increase as the load increases. Wear 

is mild and tends to be corrosive to the left of 0 and adhesive to the right 

of B. The location of B is quite variable and depends on the corrosivity 

of the lubricant formulation. For a noncorrosive lubricant, adhesive wear can 

occur at X. On the other hand, a corrosive additive can extend the boundary 

regime to Z1 before boundary film failure occurs. 

24 
Reqion YZ. - This is the regime of severe wear where scuffing and scoring 

occur. Machinery cannot operate successfully in this region and, therefore, 

the location of this transitlon point is quite important. At point Z, there 

is total surface failure and seizure occurs. 

Wear Particles and Surface Damage 

The types of wear particles and surface damage generated during these 

lubrication regimes has been reported by Reda, et al. (ref. 9). On that study, 

wear particles were generated in sliding steel contacts and were isolated by 

Ferrographic analysis (ref. 10). Basically, this technique involves pumping a 

quantlty of used lubricant across a glass slide which sits on top of an 

electromagnet (fig. 3). The ferromagnetic wear particles are magnetically 

precipitated onto the slide and can then be observed microscopically. Surface 

damage can also be observed by microscopic examination. 

Hydrodynamic and EHD regimes. - Since the surfaces are completely separated 

in these regimes, no wear or surface damage should be evident. This is the 

case with a featureless surface (fig. 4(a)) and only a few isolated wear 

partlcles (fig. 4(b)). 

Hixed and boundary lubrication regimes. - These are mild wear regimes where 

penetratqon of the boundary film occurs. 

This produces the surface damage as i1,lustrated in figure 5(a). Wear 

particles are generated in a thin surface layer that Is continuously removed 

and reformed during the sliding process. In Fertographic terminology the wear 

particles generated in these regimes are referred to as normal rubbing wear 

particles. These flake-like particles are released into the lubricant by an 

exfoliation or fatigue-like process. The rate of removal of this surface layer 

is less than its rate of formation. Wear occurs continuously but at a low rate. 

These wear particles are arranged In strings by the magnetic field of the 

Ferrograph (fig. 5(b)). A scanning electron micrograph (fig. 6) (ref. 11) at 

a higher magnification Illustrates their flake-like nature. Typically, for 

steel surfaces, these particles are 0.75 to 1.0 pm in thickness wlth a major 

dimension of less than 15 pm. 

The transition from the EHD regime into the mixed and boundary regimes is 

dramatically illustrated by Ferrographic analysis (fig. 7) (ref. 12). In this 

figure the Ferrogram density is plotted as a function of the A ratio. 

Ferrogram density is a measure of the amount of wear particles. The A 

ratio is the ratio of the film thickness to the composite surface roughness. 

This data was generated by sliding steel balls of three different roughnesses 

against a sapphire plate. As can be seen, the amount of wear debris increases 

sharply at A values of one and below where surface interactions begin taking place. The comparable curve for the friction coefficient (f) as a function 

of the A ratio appear in figure 8 (ref. 12). Here the corresponding 

increase in the friction with increasing surface interactions is evident. This 

is analogous to the rising portion of the Stribeck-Hersey curve (fig. 1) in the 

mixed film regime. 

Transition to severe wear. - As load is increased, there is some point (Y) 

(fig. 2) where the rate of removal of the surface layer starts to exceed its 

25 
rate of formation. A transition from mild to severe wear occurs. Surface 

damage becomes more extenslve (fig. 9(a)). As the surface film starts to fall, 

much larger metallic wear particles (up to 150 pm in major dimension) are 

formed (fig. 9(b)). 

Mild oxidative wear. - Under certain conditions a mild form of oxidative 

wear occurs. 

type (hematite). The surfaces appear oxidized and .grooved to depths of 20 urn 

(fig. 10(a)). The wear particles appear reddish-orange In reflected white 

light (fig. 10(b)). 

Here the majority of wear particles are iron oxide of the aFe203 

Severe oxidative wear. - As load is further increased, a transition from 

mild to severe oxidatlve wear occurs. The surfaces are now grooved to depths 

of 100 vm and black oxide appears on the surface (fig. ll(a)). The generation 

of black oxide particles consisting of y-Fe20g, Fe304 and FeO predominates 

(fig. ll(b)). 

Both of these oxidative wear regimes are most commonly observed durlng 

unlubricated conditions. However, both regimes have been observed during lubricated sliding which is Indicative of very poor, perhaps starved, lubrication. 

Seizure. - Finally, at very high loads, complete breakdown of the wear 

surface occurs. Considerable smearing and tearing of the surface is evident 

with grooving to 200 urn (fig. 12(a)): Free metal1;c partlcles havjng dimensions up to 1 m are generated (fig. 12(b)). 

Boundary Film Formation 

As discussed in the introduction, the most important aspect of boundary 

lubrication is the formation of surface films (from additives or the lubrlcant 

itself) which will protect the contacting surfaces. Basically, there are three 

mechanisms of boundary film formation: physical adsorption, chemlsorption and 

chemical reaction (ref. 13). 

Physical adsorption. - Physical adsorption involves intermolecular forces 

analogous to those involved in condensation of vapors to liquids. Physical 

adsorption is usually rapid, reversible, and nonspecific. Energies involved 

in physical adsorption are In the range of heats of condensation. Physical 

adsorption may be monomolecular or multilayer. 

In this process. An Idealized example of the physical adsorption of hexadecanol on an unreactive metal is shown In figure 13. Because of the weak 

bonding energies involved, physically adsorbed species are not very effective 

boundary lubricants. 

There is no electron transfer 

Chemisorption. - Chemisorption of,a specles on a surface is usually specific, may be rapid or slow and is not always reversible. Energies involved 

in chemisorption are large enough to imply that a chemical bond has formed 

(i.e., electron transfer has taken place). Chemisorption Is a monomolecular 

process. It also may require an actlvatlon energy as opposed to physical 

adsorption which requires none. A species may be physically adsorbed at low 

temperatures and chemisorbed at hlgher temperatures. In addltion, physical 

adsorptlon may occur on top of a chemisorbed fllm. An example of the chemisorption of stearic acid on an iron oxlde surface to form iron stearate 1s 

illustrated in figure 14. 

26 
Chemical reaction. -. Although chemisorbed films involves a chemical reaction, this section mainly deals with inorganic reaction products on surfaces. 

This process is also specific, may be rapid or slow (depending on temperature, 

reactivity and other conditions), and is irreversible. Films can be unlimited 

in thickness. An idealized example of a reacted fllm of iron sulfide on an 

Iron surface is shown in figure 15. 

Another chemical reaction of import to boundary lubrication is "frtction 

polymer" formatlon. In 1958, Hermance and Egan (ref. 14) reported on the 

occurrence of organic deposits on electrical relay contacts and coined the term 

"friction polymer." Since that time many investigators have observed the presence of this material in lubricated contacts. Although little Is known about 

its mode of formation or Its chemical structure, it can, in some cases, act as 

a boundary lubricant (ref. 14). An example of "friction polymer" debris generated by a polyphenyl ether appears in figure 16 (ref. 15). A summary of the 

importance of friction polymers In lubricated contacts appears in reference 16. 

Physical Properties of Boundary Fllms 

The physical properties of boundary films that are Important in determining 

their effectiveness in protecting surfaces include: melting or decomposition 

temperature, shear strength, thickness, surface adhesion, cohesion and solubility In the bulk lubricant. 

Heltinq point. - The melting point of boundary films is probably the most 

common property which correlates with film failure. The literature is replete 

with such examples. Russell, et al., (ref. 17) reported friction transitions 

for copper lubricated with pure hydrocarbons. Friction data for two hydrocarbons (mesitylenes and dotrlacontane) appear In figure 17 as a function of temperature. Here boundary film failure occurs at the melting point of each 

hydrocarbon. Although data for only two hydrocarbons are shown in figure 17, 

the same phenomenon occurred for several other compounds. Obviously, the films 

of these nonpolar materials, which are not chemically bound to the copper surface, provide little protection in the liquid state. 

In contrast, chemisorption of fatty acids on reactive metals ylelds fallure 

temperature based on the softening point of the soap rather than the melting 

point of the parent fatty acid. Examples of transition temperatures for several fatty acids appear in figure 18 (ref. 18). 

Chemically reacted Inorganic surface films such as oxides and sulfiues do 

not have failure transitions that correlate with their melting points. These 

materials often have very high melting points (>lo00 "C) and other factors 

(such as decomposition or physical removal) may limit their effectiveness at 

temperatures well below their melting point. 

Shear strength. - The shear strength of a boundary lubricating film should 

be directly reflected in the friction coefficient. In general, this is true 

with low shear strength soaps yielding low friction while hlgh shear strength 

salts yield high friction (ref. 5). However, the important parameter in boundary friction is the ratio of the shear strength of the fllm to that of the substrate (ref. 5). This relationshlp is illustrated in figure 19. Shear strength 

is also affected by both pressure and temperature. For example, shear stress 

as a function of load (pressure) for stearlc acid is shown In figure 20 

27 
(ref. 19). A compilation of similar data for inorganic compounds appears in 

reference 20. 

Boundary-film thickness. - boundary film thicknesses can vary from a few 

angstroms (adsorbed gas) to thousands of angstroms (chemical reactlon film). 

The effect of thickness on fricti'on has been discussed In detail by Kragelskii 

(ref. 21). In general, as the thickness of a boundary film increases, the 

coefficien-t of friction decreases (fig. 21(a)). However, continued lncreases 

in thickness may result in an Increase in friction (fig. 2l(b)). This figure 

has the general form of the familiar Stribeck-Hersey curve (fig. 1). Another 

point of interest: shear strength of all boundary films decreases as their 

thickness Increases, which may be related to the above. 

For physically adsorbed or chemisorbed films, surface protectlon is usually 

enhanced by Increasing film thickness (ref. 18). The frictional transition 

temperature of multilayers also Increases wi,th increasing number of layers 

(ref. 3). - 

For thick chemically reacted films there is an optimum thickness for minimum wear which depends on conditions. The relationship between wear and lubricant (or addttive) reactivity is shown in figure 22. Here, if reactivity is 

not great enough to produce a thtck enough film, adhesive wear occurs. On the 

other hand, if the material is too reactlve, very thick films are formed and 

corrosive wear ensues. Lubricant or additive reactivity is also a function-of 

temperature Cflg. 23) (ref. 22) and concentration (fig. 24) 

(ref. 7). 

Effect of Operating Variables on Friction 

- Load. 

(fig. 1). 

essential 

- As mentioned In the earlier discussion on the Stribeck-Hersey curve 

in the boundary lubrication regime, the coefficient of friction is 

ly constant with increasing load. This is a statement of Amonton's 

law which says that the coefficient of frictlon is independent of load. This 

law is amazingly well obeyed for most systems if there is no boundary film 

failure. This is illustrated in figure 25 (ref. 23) for copper surfaces lubricated with two fatty acids. At loads (>SO g) the coefficientlent of friction 

is essentially constant. The Increasing friction at decreasing loads is probably related to molecular orientation effects and the fact that film penetration does not occur. 

Speed. - In general, in the absence of viscgsity effects, friction changes 

little with speed over a sliding speed range of 0.005 1.0 cm/sec (ref. 4). 

Where viscosity effects do come Into play, two types of behavior are observed. 

These are Illustrated in figure 26 (ref. 24). Here, relatively nonpolar materials such as mineral oils show a decrease In friction with increasing speed 

while polar fatty acids show the opposite trend. In addition, the mineral oil 

behavior ls usually assoclated with stlck-slip phenomena. This is of practlcal 

importance since a good boundary lubricant does not lead to stick-slip behavior. 

At high speeds, viscous effects will be present and increases in frlction 

are normally observed (fig. 27) (ref. 25). This portion of the frictlon curve 

is analgous to the rising portion of the Stribeck-Hersey curve (fig. 1) as one 

approaches hydrodynamic lubrication. 

28 
Temperature. - It is impossible to generallze about the effect of temperature on boundary friction. So much depends on the other conditions and the 

type of materials present. Temperature can cause disruption, desorption, or 

decomposition of boundary films. It can also provide activatelon energy for 

chemisorption or chemical reactions. The frictional transition temperatures 

of figure 18 show the effect of temperature on the melting of the chemisorbed 

fatty acids. 

Atmosphere. - The presence of oxygen and water vapor in the atmosphere can 

greatly affect the chemical processes that occur in the boundary layer. These 

processes can, In turn, affect the friction coefficient. The importance of 

atmospheric and adsorbed oxygen Is illustrated in figure 28 (ref. 26). Here, 

the "EP" activjty of tricresylphosphate (TCP) is totally absent in a dry nitrogen atmosphere. In contrast, normal t'EP" activity is present in dry air. 

Effect of Operating Variables on Wear 

- Load. - It is generally agreed that wear increases with Increasing load but 

no simple relationship seems to exist. This refers to the situation where no 

transition to severe wear has occurred. At this point, a discontinuity on wear 

versus load occurs which Is il.lustrated in figure 2. Figure 29 Illustrates the 

increase in wear scar radlus with increasing load (ref. 4). 

Speed. - For practical purposes, wear rate In the boundary lubrication 

regime Is essentially independent of speed. Obviously, this does not hold If 

one moves Into the EHD regime with Increasing speed. This also assumes no 

boundary film failure due to contact temperature rise. 

as a function of sliding speed appears in figure 30 (ref. 28). These data 

indicate that fluid film effects are negligible only at 25 and 50 rpm and a 

1 kg load. 

An example of wear rate 

Temperature. - As was the cise for friction, there is no way to generalize 

the effect of temperature on wear. 

tlon also pertain to wear. 

is presented in figure 23. 

The same statements that pertain to fricA good example of the effect of temperature on wear 

Atmosphere. - The effects of atmospheric oxygen and moisture on wear has 

been studieo by many investigators. Oxygen has been shown to be an impcrtant 

ingredient in boundary lubrication experiments involving load carrying additives (refs. 4, and 29 to 33). For example, the presence of oxygen or moisture 

in the test atmosphere has a great effect on the wear properties of lubricants 

containing aromatic species (figs. 31 and 32) (refs. 27 and 33). 

Additive Behavior of Boundary Lubricated Systems 

!n discussing figure 2, it was stated that the boundary lubrication reglme 

could be extended to higher loads by proper formulation of the lubricant with 

additives. A discussion of the variety of boundary additives is outside the 

scope of this paper. However, the general behavior of the two common types of 

boundary addltives, namely, antiwear and "EP", is illustrated in figure 33 

(ref. 34). In figure 33 wear rate (K) Is plotted as a function of load (F). 

There is a wear transition when boundary film failure occurs. The presence of 

an antiwear additive reduces wear (AK) but may have little or no effect on 

29 
the wear transition load. On the other hand, an "EP" additive yields an 

increase In the load carrying capacity (AF) with little or no effect on the 

wear rate below the orlglnal base oil wear transition. 

The most comnon antlwear additlves are those which contaln the element 

phosphorus. Typical examples lnclude the metal dlalkyldithlophosphates, 

organlc phosphates and phosphites. "EP" addltlves lnclude compounds containing 

sulfur and chlorine. A review of these two additlve classes appears in 

reference 35. 

CONCLUDING REMARKS 

In conclusion, the boundary lubrication regime has been shown to be a complex arena of a variety of competing chemical and physical processes. 

ultimate unuerstanding of this regime will come when these processes and their 

interrelationships are completely identifled. 

The 

REFERENCES 

1. 

2. 

3. 

4. 

5. 

6. 

7. 

8. 

9. 

Stribeck, R.: "Characterlstlcs of Plain and Roller Bearlngs", m. 

V.D.I., Vol. 46 (1902). . 

Hersey, H.D. : 

3. Wash. Acad. Scl., 4 1914, pp. 542-52. 

Godfrey, U.: 

Understanding Boundary Lubrication," Fundamentals of Tribology, ed. by N. 

P. Suh and N. Saka, MIT Press (1980), pp. 945-967. 

"The Laws of Lubrication of Horizontal Journal Bearings", 

"Review of Usefulness of New Surface Analysis Instruments In 

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Appraisal of the World Literature. American Society of Mechanical 

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Godfrey, 0.: "Boundary Lubricatton," in Interdlsclpllnary Approach to 

Friction and Wear, NASA SP-181, ed. by P.M. Ku (1968), p. 335-384. 

Feln. R. S.. "Chemlstry in Concentrated-Conjunction Lubrlcation," Interdlscipl inary Approach to the Lubrlcatlon of Concentrated Contacts, NASA 

SP-237, ed. by P.M. Ku (1970), p. 489-527. 

Rowe, C.N.: "Wear-Corrosion and Eroslon," Interdlsclpllnary Approach to 

Llquid Lubricant Technoloqy, NASA SP-318, ed. by P.M. Ku (1973). 

p. 527-568. 

Beerbower, A.: 

tions Forcasts, Department of the Army, DAHC-19-69-C-0033 (1972). 

Reda, A.A., Bowden, R., and Westcott, V.C.: 'Characterlstlcs of Particles 

Generated at the Interface Between Sliding Steel Surfaces," Wear, 34, 

"Boundary Lubrication," Scientlflc and Technlcal Appllcapp. 261-273 (1975). 

10. Seifert, W.W. and Westcott, V.C.: "A Method for the Study of Wear Particles in Lubrlcating 011," Wear, 21, 27-42 (1972). 

30 
11. Jones, W.R. Jr.: 

12. Jones, W.R., Jr., Nagaraj, H.S., and Winer, W.O.: "Ferrographic Analysis 

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Trans., 21 (3), 181-190 (1978). 

13. Adamson, A.W.:. Physical Chemistry of Surfaces, John Wiley and Sons, 2nd. 

ed. (1967). 

14. Hermance, H.W. and Egan, T.F., "Organic Deposits on Precious Metal Contacts," 8ell Syst Tech. J., 37, 739-76 (1958). 

15. Jones, W.R., Jr., nFerrographic Analysis of Wear Debris from boundary 

Lubrication Experiments with a Five Ring Polyphenyl Ether," ASLE Trans., 

18 (3). 153-62 (July 1975). 

16. Lauer, James L.; Jones, William R., Jr.; and Wedeven, Lavern 0.: Assess17. Russell, J.A., Campbell, W.E., Burton, R.A., and Ku, P.M.: "Boundary Lubment of Lubricated Contacts - Friction Polymer, Proposed NASA TM. 

rication Behavior of Organic Films at Low Temperatures," ASLE Trans., 8(1), 

48 (1965). 

18. Bowden, F.P., and Tabor, 0.: The Friction and Lubrication of Solids. Part 

19. Akhmatou, A.S.: Molecular Physics of Boundary Friction, Israel Program for 

- 1, Oxford University Press (1971) p. 182. 

Scientiflc Translations (1966). p. 248. 

20. Bridgman, P.W. : "Shearing Phenomena' at High Pressures, Particularly in 

Inorganic Compounds", Proceedings of the Amer. Acau. of Arts and Sciences, 

71 (9), 387-460 (1937). 

21. Kragelski, I.V.: Friction and Wear, Butterworths (1965). pp. 158-163. 

22. Jones, William R., Jr.: Boundary Lubricatlon of Formulated C-Ethers in Air 

to 300 OC," Lubr. Eng., 32, (lo), 530-538 (1976). 

23. Whitehead, J.R.: "Surface deformation and Friction of Metals at Light 

Loads," Proc. R. SOC. London Ser. A, 201 (1064) 109-124, (1950). 

24. Clayton, 0.: An Introduction to Boundary and Extreme Pressure Lubrication. 

Physics of Lubrication, Br. J. Appl. Phys. 2, Suppl. 1, 1951, p. 25. 

?5. Forrester, P.G.: "The Influence of Sliding Velocity and Other Variables 

on Kinetic Frictlon In or near the Boundary Region," Proc. R. SOC. London, 

Ser. A, 187 (loll), 439-463 (1946). 

26. Faut, O.D.; and Wheeler, 0.: On the Mechanism of Lubrication by Tricresyl 

Phosphate (TCP). I. The Coefficient of Friction as a Function of Temperature for TCP on M-50 Steel, NASA TP in process. 

31 
27. Jones, W.R.., 3r. and Hady, 'W.F.: "Effect of Humidlty and a Wettability 

Additlve on Polyphenyl Ether Boundary Lubrication of Steel In Air and 

Nitrogen to 350 OC," NASA TN 0-6055 (1970). 

28. Loomls, W.R., and Jones, W.R.., Jr.: "Steady-State Wear and Friction in 

Boundary Lubricatlon Studies," NASA TP-1658 (1980). 

29. Godfrey, D., 

30. Goldblatt, I.L., and Appeldorn, J. K.: The Antiwear Behavior of TCP In 

"The Lubricatiopn Hechanlsm of Trlcresyl Phosphate on Steel ,It 

ASLE Trans. 8 (l), 1-11 (1965)." 

Different Atmospheres and Different Base Stocks," ASLE Trans. 13, (3), 

203-21 4 ( 1970). 

31. Wheeler, D.R., "Surface Chemistry of Iron Sliding in Air and Nitrogen 

Lubricated With Hexadecane and Hexadecane Containlng Dlbenzyl-Disulfide," 

NASA TP-1545 (1979). 

32. Vlnogradov, G.V., Arkarova, V. V., and Petrov, A. A.: "Antiwear and 

Antifrictionlon Propertles of Hydrocarbons Under Heavy Loads", Wear, 4, 

(4) 274-291 (1961). 

33. Jones, W.R., Jr.: "The Effect of Oxygen Concentration on the BoundaryLubricating Characteristics of a C-Ether and a Polyphenyl Ether to 300-°C,t8 

Ws, 73, 123-136 (1981). 

34. Czichos, H.: "Tribology, a Systems Approach to the Science and Technology 

of Friction,n Lubrication and Wear, Elsevler (1978). 

35. Forbes, E.S.: "Antiwear and Extreme Pressure Additives for Lubricants," 

Trlbolow, 3, 145-152 (1970). 

32 
-6 -4 

.. h- 10 - 10 IN 

-1 

h- 10 IN I -- 

.E 

.001 

4" 

L 

--- 

I I I 

I I. 

ZN (VISCOSITY) (VELOCITY) 

Figure L - Coefficient of friction as a function of speebvelocit).c 

(LOAD) -I 

P 

load parameter (Stribeck-Hersey cunrel.(ref. 1). 

SEIZURE 

WEA 

f I ZI TRANSITION I 

0- 

/' 

0' 

BOUN DA RY 

I 

AX B Y 2 

i HYDRO- RELATIVE LOAD 

DYNAMIC 

LUBRICATION 

Figure 2. - Wear rate as a function of relative load depicting the various regimes of lubrication (ref. 8). 

33 
OIL 

OIL SAMPLE7 

INLET- / 

-ENTRY POSITION 

Figure 3. - Ferrugraph analyzer. 

34 
r; 

ORIGINAL PACE 

BLACK AND WHITE PHOTOG3WM 

1 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 4. - Hydrodynamic or elastohydrodynamic 

I ubricat ion regime. 

35 
ORiGfNAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 5. - Boundary lubrication regime. 

36 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

Figure 6. - Normal rubbing wear particles (from 

ref. 11). 

30 F 

20 

10 

0 SMOOTH BALL 

A MEDIUM ROUGH BALL 

0 ROUGH BALL 

I - 0 4 n 12 16 20 24 

FILM THlCKNESSlCOMPOS ITE SURFACE ROUGHNESS, 

havg 

Figure 7. - Ferrogram densitv versus average A ratio 

(ref. 12). 

37 
0 4 8 12 16 20 24 .os 

FILM THICKNESS/COMPOSITE SURFACE ROUGHNESS, Aaq 

Figure 8. - Friction versus average A ratio (ref. 

12). 

38 
8 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 9. - Transition to severe wear. 

39 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

--- 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 10. - Severe wear regime-mild oxidative 

wear. 

40 
0 si G i r4.4 I- P,4 G E 

BLACK AND WhlTE PtiOTOGRAPH 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 11. - Severe wear regime-severe oxidative 

wear. 

41 
ORlGlNAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

'? .- -. -,e--- 

(a) Wear surface. 

(b) Wear particles (ref. 9). 

Figure 12. - Transition to seizure. 

42 
z .- 

c 

m 

c 

0 

z =! IA 

I 

43 
Figure 15. - Inorganic film formed by reaction of 

sulfur with iron to form iron sulfide (ref. 5). 

Figure 16. - Photomicrograph of the wear debris 

generated by a polyphenyl ether at 100' C in 

dry test atmosphere of 1 percent oxygen and 

99 percent nitrogen (by volume)(ref. 15). 

44 
1" 

1" 

45 
SF SHEAR STRENGTH OF FILM 

- SHEAR STRENGTH OF MnAL sm 

I I 1 I I I I 

.5 

'F'* m 

Figure 19. - Friction as a function of the ratio of shear strengths 

of film and metal (ref. 5). 

0 

Figure 20. - Shear stress as a function of load for a thin layer 

of solld stearic acid (ref. 19). 

46 
1.2 I 

1. 1 

1.0 

0.9 

0.8 

0.7 

L 

0 

5 

El 

u o( 

0.4 L 

0.3 

0.2 

0. 1 

0 

1 2 3 4 5 

EQUIVALENT OXIDE FILM THICKNESS, lF5 cm 

(a) 

c 

10-6 10-5 10-4 10-3 3.10-3 

FILM THICKNESS, cm 

(b) 

Figure 21. - Relationship between friction and thickness 

of surface films: (a) coefficient of friction against oxide 

film thickness on copper; (b) coefficient of friction 

against thickness of indium film (ref. 21). 

47 
.- c 

E 

m' 

E 

z 

TEST CONDITION 

B 

1 ____) 

ADHESIVE WEAR CORROSIVE WEAR 

~~ 

LUBRICANT OR ADDITIVE REACTIVITY - 

Figure 22. - Relationship between wear and lubricant 

reactivity (ref. 7). 

10-13 BASE FLUID 

-0- FORMULATION I 

----- FORMULATION II / 

TEMPERATURE, OC 

Figure 23. - Wear versus temperature for various C-ether 

formulations (ref. 22). 

48 
E 

E 

.6 - 

.5 ,r TRICRESYL PHOSPHATE 

.4 

.3 

.2- 

- 

- 

- 

.1 I I 

.ooo1 .all .01 .1 1 10 

CONCENTRATION OF PHOSPHORUS, wt % 

Figure 24. - Wear versus additive concentration fmm foup ball tests (ref. 7). 

-..-- OCTACOSANOIC ACID 

*75b\,, 

; 

MURK ACID 

101 102 103 104 

0 

1r2 10-1 100 

LOAD, g 

Figure 25. - Friction versus load for copper lubricated with 

fatty acids (ref. 23). 

49 
* I 

.12 

.08 

.04 

u MINERAL OILS 

- 

- 

Ah 1 

0 .1 .2 .3 .4 

SPEED, crnlsec 

Figure 26. - Friction versus speed (ref. 24). 

SPEED, cm/sec 

Figure 27. - Friction versus speed for steel5ronze Iubricated with oleic acid (ref. 25). 

50 
L 

500 

400 

300 

200 

100 

F DRY NITROGEN 

I 

I 

I 

I 

I 

I 

0 50 100 150 200 250 300 350 400 

TEMPERATURE, OC 

Figure 28. - Friction versus temperature for pure tricresyl 

phosphate on M-50 steel. (Ref. 26) 

12 

0 360 720 1080 

TIME, min 

Figure 29. Radius of wear scar as a 

function of time for several loads. 

(Ref. 4) 

57 
d 

w 

w 

CL 

VI 

Y !c 

n 
L. 

E 

E 

cc\ 

d 

L 

a 

a 

e 

W 

0 DRY NITROGEN 

WRNITROGEN 

n DRYAIR 

W€r AIR 

T? ._ .. . ,. . .. , :: 

EO 350 

DISK TEMPERATURE, OC 

Figure 32. - Rider wear for polyphenyl ether in wet and 

dry nitrogen and in wet and dry air at disk temperatures 

of 1500, 2500, and 3W C. (Ref. 27) 

LOAD, F 

BASE OIL?, 

\ 

\ 

EP ADDITIVE-, 

'\ 

Figure 33. - Wear behavior of boundary-lubrication systems (ref. 

34). 

53 
SOLID LUBRICANT MATERIALS FOR HIGH TEMPERATURES - A REVIEW* 

Harold E. Sliney 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

Solid lubricants that can be used above 300 "C in air are discussed. The 

scope of the paper includes coatings and self-lubricating composite bearing 

materials. The lubricants considered are r e p r e s e n t a t i v e d i c h a l c o g e n i d e s , 

graphite, graphite fluoride, polyimides, soft oxides, oxidatively stable 

fluorides, and hard coating materials. A few general design considerations 

revelant to solid lubrication are interspersed throughout the paper. 

INTRODUCTION 

This paper is a review of selected research data which illustrate the tribological properties of materials that are likely to be or are now in use'as 

high-temperature solid lubricants. The term "high temperature" is relative to 

one's frame of reference; therefore a definition is necessary. For this paper 

we will somewhat arbitrarily define "high temperature" solid lubricants as 

those that will not rapidly oxidize or otherwise thermally degrade in air at 

temperatures to at least 300 "C. Above this temperature most oils, greases, 

and all but a few organic polymers are not oxidatively stable for any appreciable time. 

The scope of this paper includes coatings and self-lubricating composites. 

The materials considered are r e p r e s e n t a t i v e d i c h a l c o g e n i d e s , graphite, graphite 

fluoride, polyimides, soft oxides, oxidatively stable fluorides, and hard coating materials. A few general design considerations relevant to solid lubrication are interspersed throughout the paper. 

LAYER LATTICE SOLID LUBRICANTS 

Layer lattice solid lubricants have a hexagonal-layered crystal structure. 

Their shear properties are anisotropic with preferred planes for easy shear 

parallel to the basal planes of the crystallites. In some of the compounds 

such as molybdenum disulfide (MoSz), a low shear strength is intrinsic to the 

pure material. In others, notably graphite, the presence of absorbed gases or 

intercalated "impurities" between the basal planes appears to be necessary to 

develop desirable friction characteristics (ref. 1). 

The most common representatives of this class of lubricants are graphite 

and the dichalcogenides, notably MoS2 and WS2. Since a large volume of literature exists for these materials (e.g., ref. 2), they will not be discussed in 

detail here except to emphasize the importance of chemical reactivity in determining the suitability of a solid lubricant for a specific applicatlon. The 

maximum useful temperatures for solid lubricants depends strongly on the composition of the ambient atmosphere, the required life at temperature, and factors 

*Prepared for the Special Issue "Dry Bearings" of the Tribology 

International Journal, October 1982. 

55 

FRECEDfNG PAGE BIAXK NOT FILMED 
such as oxygen availability at the lubricated surface (is the coating openly 

exposed to the atmosphere or shielded within conforming bearing surfaces), airflow rates, lubricant particle size, and the influence of adjuvants and 

binders. 

The Dichalcogenides 

The maximum temperature for lubrication with MoS2 In an air atmosphere is 

limited by oxidation to about 400 "C under favorable conditions. Some oxidation kinetics data for loosely compacted MoS2 powders of 1-pm average particle 

slze are given in figure l(a) (from ref. 3). At a modest airflow rate over the 

compact, 50 percent of the MoS2 was oxidized to molybdic oxide (MoO3) in 1 hour 

at 400 "C. At a six times higher airflow rate, the temperature for an oxtdation half-life of 1 hour was reduced to 300 "C. Figure l(b) compares the oxidation kinetics of MoS2 and WS2 at the lower airflow rate. It is interesting 

that the curves for MoS2 and WS2 intersect, with MoS2 oxidizing more rapidly 

above about 340 "C. Friction experiments were conducted with a pin-on-disk 

apparatus using a hemispherically tipped pin in sliding contact with the flat 

surface of a rotating disk. A comparison of the oxidation data of figure 1 and 

the friction data of figure 2(b) (from ref. 3) shows that the loss of lubricating abllity of MoS2 and WS2 in alr coincides with the temperatures at which 

rapid conversion to the oxides occurs. 

Figure 2(a) also shows that both compounds lubricate to much higher temperatures in a nonreactive argon atmosphere. In an Inert gas or vacuum the maximum useful temperature Is a function of the thermal dissociation rates rather 

than the oxidation rates of the lubricants. Thermal dissociation rates and the 

friction coefficients of molybdenum and tungsten disulfides, diselenides, and 

ditellurldes in vacuum have been systematically studied (ref. 4). The major 

results, summarized in table I, indicate that the dlsulfides are the most 

stable, the diselenides are intermediate, and the dltellurides are the least 

stable. However, thin burnished films of the diselenides with their higher 

densities evaporate more slowly than the disulfides. Apparently for the very 

thin, burnished films used in this study, the evaporation rates were the controlllng factor in determining the maximum temperature for effective lubrlcation. A more recent paper described the thermal stability and lubricating 

characteristics of various bonded MoS2 coatings in vacuum (ref. 5). The 

limiting temperatures for a significant wear life of these coatings ranged from 

600 to 700 "C in agreement with the 650 "C limiting temperature for burnished 

MoS2 reported in reference 4. 

Graphite 

Although MoS2 and WS2 are "intrinsic" solid lubricants that are most effective In a vacuum or a nonreactive gas atmosphere, It has been long known (ref. 6) 

that graphite, in contrast, is a poor jntrinsic lubricant that requires the 

presence of absorbed vapors such as water or hydrocarbons to develop good lubricating properties. Certain solid adjuvants such as cadmium oxide (CdO) and 

other oxides or salts are also helpful in improving the lubricating abillty of 

graphite. Figure 3 (from ref. 7) illustrates the remarkable effect of CdO 

addition on l u b r i c a t i o n w i t h graphite powder from room temperature to 540 "C. 

Without the CdO addition graphite lubricated at room temperature and above 

56 
425 "C but not at intermediate temperatures. Low frlction at room temperature 

was attributed to the beneficial effect of absorbed moisture. High friction 

at intermediate temperatures was attributed to desorption of water and possibly 

other gases. Low friction above 425 OC was attributed to interactlon of graphite with oxides of the lubricated metal. Graphite Itself begins to oxldize at 

about 400 OC. Nevertheless it has been reported that graphite can be used as a 

wqre-drawing lubricant for tungsten and molybdenum at temperatures as high as 

1100 OC in ambient air conditions (ref. 8). Oxidation of graphite and of the 

metals occurs rapidly at the drawing temperatures. However, the time at high 

temperature is so short that effective lubrication is achieved. Therefore 

high-temperature lubrication must take into account the interaction of the 

lubricant, the atmosphere, and the surface being lubricated and must further 

balance reaction rates against the tequlred residence tlme of the lubricant 

within the contact. 

Graphite Fluoride 

Graphite fluoride (CFx)n, also referred to as carbon monofluoride (when 

x = 1) is a relatively new solid lubrlcant that can be loosely described as a 

layer lattice intercalation compound of graphite. It is prepared by the direct 

reaction of graphite with fluorine gas at controlled temperature and pressure. 

It is gray to pure white depending on its stoichiometry. The subscript ICx" in 

(CFX), can vary from about 0.3 to 1.1. For x 2 1, the compound Is pure whlte, 

electrically nonconductive, and nonwettable by water (hydrophobic). There is 

some debate about whether (CFX), is a true intercalation compound because 

the basal planes of the graphite crystallites are distorted to a puckered, nonplanar configuration when the compound is formed. However, there is no doubt 

that the original graphite crystal lattlce is the primordial lattice from which 

the crystal structure of (CF,), is formed. 

covalent, with the fluorine atoms located between the distorted basal planes. 

The spacing between the basal planes is expanded from 3.4 A in graphite to 

7.521.5 A in (CFx)n (ref. 9). The crystal lattice expansion and the distortion of the carbon basal planes are schematically represented and compared 

with the crystal structure of graphlte In figure 4. Although (CFx)n is not 

known to oxidize in air, it decomposes thermally above about 450 OC to form 

carbon tetrafluoromethane, other low-molecular-weight fluorocarbons, and carbon 

(ref. 10). 

The fluorine-to-carbon bonds are 

Some early research on the lubricating properties of (CFX), was reported In 

reference 11. In this study, thln lubricating films of (CFX), were burnlshed 

on 440C and 301 stainless steel disks and evaluated in pin-on-disk experlments. 

Wear life and friction coefficient data for burnished (CFx)n and MoS2 films on 

440C are compared in figure 5. Film failure was considered to be the time at 

which the friction coefficient exceeded 0.3. The (CFX), films were the more 

durable over the entlre temperature range shown. Friction coefficients were 

for the most part well below 0.1 up to the fallure temperatures of the coatings, which are indlcated by the arrows in figure 5. Failure temperature was 

400 "C for MoS2 and 480 OC for (CF,),. These failure temperatures correlate 

well with expectations based on the oxidation kinetics data for MoS2 (fig. 1) 

and the previously referenced thermal decomposition temperature of (CFX),. 

Similar results were obtained for the lubrlcatlon of 301 stainless steel with 

(CFX),, but burnished MoS2 fllms falled immediately on this alloy. 

agreement wlth these results was obtained wlth burnished (CFX), by uslng a flat 

rub block on a cylinder specimen conflguratlon (ref. 12). 

General 

57 
(CFX), exhibits an extreme degree of plasticity within a lubricated contact. It has been observed (ref. 13) that all this ability to readily undergo 

plastic flow within the contact Is a characteristic that all good solid lubricants have in common; nonlubricated solid particles tend to fragment within the 

contact and, if of sufficient hardness, will embed In one surface and abrade 

the other. To illustrate the two extremes, compare the plastic behavior of 

(CFX), shown in figure 6 with the brittle behavior of SIC in figure 7. MoS2 

and graphlte also readily flow within a moving concentrated contact. 

order of plasticity subjectively observed is (CFX), > MoS2 > graphlte. 

The 

Various resin binders have been used with (CFX), to achieve longer wear 

lives and higher load capacities than can be achieved with burnished films. 

Good results were generally obtained in regard to low friction and wear, but 

there is some discrepancy in the reported load capacity of the coatings in 

Falex V-block tests. Reference 14 for example reported a high load capacity 

for epoxy - phenolic resin - bonder (CFx)n, but the authors of reference 15 

reported, on the basis of their Falex evaluation, that (CFx)n is not considered 

to be suitable for use in heavy load application. Therefore (CFX), coatings 

may not be the lubricant of choice for highly loaded, concentrated (Hertzian) 

sliding contacts. 

load capacity for lubricating bearings with conformal contacts such as plain 

spherical bearings or cylindrical bushings. 

ular has been extensively studied and will be discussed in the next sectlon. 

However, there are (CFX), coatings that have quite adequate 

Polylmide-bonded (CFX), in particPOLYIMIDES 

Polyimide Coatings 

A few organic polymers are oxidatively stable and also have glass transition temperatures above 300 "C. Examples are polyqulnoxilines, polybenzimidazoles, and polyimides. Of these, the polyimldes are by far the most readily 

available and have been the most studied by tribologists. Polyimide coatings 

have been studied as self-lubricating varnlshes (ref. 16) and as resin binders 

for inorganic solid lubricants (refs. 16.and 17). 

The results of pin-on-disk experiments with polyimide varnish coatings on 

440C stainless steel are given.in figure 8. Wear life and friction characteristics were determined in three different atmospheres: dry argon, dry air, and 

air containing lo4 ppm of water vapor. 

transition in the friction and wear life propertles of the.polyimide fllms 

between 25 and 100 "C. Above the transitlon temperature polyimide films performed well as solid lubricants. Low friction and long wear lives were 

obtained even with no solld lubricant additive In the film. At room temperature the polyimide films did not lubricate nearly as well as they did from 100 

to 500 OC. This transition in the tribological properties of polyimide has 

been attributed to second-order relaxation in the molecular bonds of the 

polymer between 25 and 100 "C (ref. 18). 

In all these atmospheres there is a 

Polyimide-Bonded Graphite Fluoride Coatings 

The effect of adding (CF,), or MoS2 to the polyimide varnish is illustrated 

in figure 9. The solid lubricant additions clearly improve the wear life and 

friction at room temperature. In effect, the undeslrable friction transition 

below 100 "C is totally masked by the addition of the solid lubricant pigments. 

58 
A promising application for polyimide-bonded (CFX), coatings is as a 

backup lubricant for compliant (foil) gas bearings at temperatures to 350 OC. 

About 100 OC higher gas temperature capability can be achieved by substituting 

PI-bonded (CFX), for the more conventional PTFE coatings. Current research 

by this author shows that these coatings are remarkably durable in start-stop 

endurance testing of foil bearings. 

Polyimide Composites, Background 

Polyimides are also used in heat-cured composite bearing materials. It is 

shown in references 19 to 21 that polyimide and polyimide compositions contalning various powder fillers had interesting possibilities as self-lubricating 

bearing materials. Polyimide bearing materials with powdered solid lubricant 

additives such as graphite and MoS2 have found many applications. However, the 

powdered additives significantly reduce the mechanical strength of the molded 

polyimide. A polyimide with a compressive strength of 207 MPa (30 000 psi) 

without additives will typically have about one-half of that compressive 

strength when 10 to 20 ut X of solid lubricant powder Is Incorporated into the 

polymer. However, lubricating properties can be improved with no loss of 

strength by using graphite fiber, which provides both lubrication and 

reinforcement. 

The potential of graphite-fiber-reinforced polyimide (GFRPI) as a selflubricating composite material was first demonstrated In accelerated crossedcylinder wear tests (ref. 22). The performance of GFRPI in oscillating plain 

spherical bearings was reported in reference 23; in that research chopped 

graphite fiber reinforcement was used. The fibers were incorporated into the 

B-staged polyimide and mixed to achieve a random fiber distribution; then the 

mixture was transfer molded into the bearing and cured under heat and pressure. 

GFRPI composite bearing materials fabricated from layers of woven graphite 

fabric and polyimide have also been reported (e.g., ref. 24). The discussion 

in this paper is confined to chopped-flber-reinforced bearing materials. 

Chopped-Fiber-Reinforced Polyimides 

Pin-on-disk studies. - The evaluatlon of chopped-GFRPI composites in a pinon-disk tribometer is reported in reference 25. In these experiments hemispherically tipped 440C stainless steel pins were slid against the flat surface of 

rotatingSGFRP1 disks. 

evaluated. The fiber properties are glven in table 11. The fiber designated 

type L is a relatively low-strength, low-modulus fiber; the type H fiber has a 

medium tensile strength and an elastic modulus approximately 10 times higher 

than the type L fiber. 

given in figure 10. Type A is an addition polyimide and is highly crosslinked. 

Type C is a condensation polyimide and has a linear, essentially noncrosslinked 

polymeric structure. Type A polyimides do not produce water vapor as a product 

Two types of graphite fiber and two polyimides were 

The molecular structures of the two polyimides are 

of the advanced 

times considered 

Figure ll(a) 

at 25 and 300 "C 

f ri ct i on coef f i c 

tage of polymerization (as do type C) and .therefore are someeasier to mold into void-free parts. 

gives the friction characteristics of the various composites 

in air containing lo4 ppm of water vapor. 

ent is about 0.2 for all of the composites, but at 300 "C the 

At 25 OC the 

59 
type AL composite has the lowest friction coefficient (- 0.05). 

traces (fig. ll(b)) of the wear tracks on the GFRPI disks after 300 000 sliding 

passes (disk revolutions) showed that the AL c o m p o s i t e s w e r e also the most 

wear resistant at both temperatures. 

Bearlnq tests. - Three self-aligning plain spherical bearing destgns and 

cylindrical bushings were tested. The spherical bearing designs shown in 

figure 12 consisted of (1) a GFRPI spherical element In a steel outer ring; and 

(2) a steel spherical element and outer ring with a 1.5-mm-thick selflubricating GFRPI liner transfer molded into the bearing and bonded to the 

outer ring. 

Profilometer 

In the first design tested, the spherical element was a molded composite 

ball of type AL GFRPI (ref. 23). Friction coefflcients for varlous fiber loadings are shown In figure 13 for temperatures from 25 to 350 OC. Oata for a 

conventional spherical bearing with a glass-fiber-reinforced PTFE liner are 

shown for comparison. Friction decreased in a regular manner with increasing 

graphite content. 

gave the lowest frlction but failed by b r i t t l e f r a c t u r e a t 3 1 5 OC and a 35-MPa 

(5000-psi) radial load. It was concluded that a fiber loading between 45 and 

60 percent Is about the optimum tradeoff between minimum torque and maximum 

dynamic load capacity. The standard PTFE-llned bearing had very low friction 

to 200 OC, but the liner extruded out of the bearing at 250 OC. 

When several modifications of the bearing designs shown in figure 12 were 

The composite ball with the highest fiber content of 60 wt x 

compared (refs. 26 and 27), it was found that the design variations had little 

influence on bearing friction (fig. 14) but had a significant effect on bearing 

load capacity (fig. 15). It is clear that much higher dynamic load capacity is 

achieved with a thin (1.5 mm) GFRPI liner between the ball and the outer ring 

than with a bearlng consisting of a GFRPI ball and a metal outer ring. Dynamic 

load capacities for cylindrical bushings with GFRPI liners are reported in 

reference 28, and they were about the s a m e a s for the lined sphericals. In 

general, GFRPI-lined oscillating plain bearings for high-load, low-speed appllcations have dynamic load capacities of about 140 HPa (20 000 psi) from room 

temperature to 260 OC and about 70 MPa (10 000 psi) at 320 "C. 

changes such as the addition of edge-retention features to prevent the liner 

from being squeezed out of the bearing at high loads are expected to further 

improve load capacity. 

Bearing deslgn 

UNCONVENTIONAL SOLID LUBRICANTS 

Soft Oxides and Fluorides, Fusion-Bonded Coatings 

In a search for even higher temperature solld lubricants much research has 

been performed on various soft oxides and wtth fluorides of alkali metals and 

alkallne earth metals. Oxides are of course obvious candidates for consideration when oxjdation-resistant compounds are required. The hard oxldes, typical 

of ceramic materials, such as alumina, sllica, and the slllcates have good wear 

resistance but generally high friction coefflcients. Furthermore unpollshed 

surfaces or wear debrls from hard oxides are abrasive to softer, metallic 

counterface materials. 

60 
On the other hand, soft oxldes such as lead monoxide (PbO) are relatively 

nonabrasive and have relatively low friction coefficients, especially at high 

temperatures, where their shear strengths are reduced to the degree that deformation occurs by plastic flow rather than brittle fracture. Binary and ternary 

eutectic oxide systems are of Interest because the melting-point suppression, 

which is the primary characteristic of eutectic systems, tends to lower the 

shear strength relative to the individual oxides. If the second oxide is a 

vitrifying agent such as Si02, glaze (glass) formation is promoted at the 

sliding surface, and this also tends to modify friction by introducing a visCOUS component of shear. This Increases or decreases friction depending on 

the viscosity of the glaze within the sliding contact. 

characteristics of an oxide coating are controlled by a mechanism involving 

either or both crystalline shear and viscous drag. 

Therefore the friction 

Increasing the surface temperature reduces both Crystalline shear strength 

and glass viscosity within the sliding contact and therefore tends to reduce 

the friction coefficient of oxide surfaces. This is illustrated In figure 16 

(ref. 29). which gives the effect of ambient temperature and sliding velocity 

on the friction coefficients of a stainless steel alloy lubricated with a coating of Pb0-4PbOeSiO . At low sliding velocitles the PbO coating lubricated 

With increasing sliding velocity frictional heating rates increased and low 

friction was achieved at ever lower ambient temperatures until at 6 m/sec friction coefficients of 0.2 or lower were observed from room temperature to 650 OC 

Because of the narrow range of temperatures at which PbO lubricates effectively 

at low velocities, Its use has been limited to high-speed, high-temperature 

applications such as the lubrication of dies for high-speed wire drawing. 

effectively over on i! y a very small temperature range of about 500 to 650 "C. 

Other studies showed that chemically stable fluorides of some Group I and 

I1 metals, such as LiF, CaF2, and BaF2, also lubricate at high temperature but 

over a broader range of temperatures than PbO. For example, coatings with compositions from the CaF2/BaF2 binary eutectic system lubricate from about 500 to 

950 "C. The tribological properties in air of fused fluoride coating with the 

cornposition 62BaF2-38CaF2 are given in figure 17 (ref. 30). Data for uncoated 

specimens are given for comparison. 

The fluorides of the rare earth metals are another group of metal halides 

that are chemically stable and have shown promise as high-temperature solid 

lubricants. In an exploratory study of their lubricating properties cerium 

trifluoride (CeF3) and lanthanum trifluoride (LaF3) were the best solid lubricants among the rare earth fluorides (ref. 31). The individual CeF3 or LaF3 

powders lubrlcated nickel-base super alloys in air to at least 1000 "C. Frlctlon coefficients were 0.3 to 0.4 from room temperature to 500 "C but averaged 

about 0.2 at higher temperature. These compounds received little further 

attention in the lubrication literature but definately should be considered 

where friction coefficients of 0.2 to 0.4 combined with good antiwear characteristics at high temperature are required. 

Coatings of oxide and fluoride compositions that melt at a lower temperatures than the substrate metal can be applied by well-known procedures for 

applying glass or porcelain enamel glazes. In brief, an aqueous slurry of the 

oxide or fluoride powders is sprayed onto the metal, cured to dryness, then 

furnace-fired above the melting point of the coating composition. 

ing, a fusion-bonded, dense coating is obtained. Good adhesion depends on 

reasonabl'y matched thermal expansion coefficients and other factors such as the 

Upon cool61 
nature of the high-temperature Interactions that take place between the melt, 

the metal, and the atmosphere during firing. Fusion-bonded fluoride coatings 

can be applied by a similar procedure. 

generally fired in air, inert or reducing atmospheres are generally used for 

fluorides to avoid contamination of the coatings w.ith oxides of the substrate 

metal. 

However, while oxide coatings are 

F 1 uo r 1 d e-Met a 1 Compo s i t e s 

Composite bearing materials in which the jolid lubricant is dispersed 

throughout the structure are advantageous when long lubricant life is required. 

In some cases a thin, bonded solld lubricant coating is used as an overlay on 

the self-lubricating composite material. This assures the minimum friction 

coefficient obtainable by enrichment of the composite surface with lubricant 

while providing long life because of the underlying, self-lubricating composite 

material. The friction and wear of a composite consisting of a porous, sintered metal matrix infiltrated with barium fluoride - calcium fluoride eutectic 

are shown in figure 18 (from ref. 32). Wear life comparisons for composite 

coatings in air and hydrogen are given in table 111. Wear life is herein 

defined as the number of sliding cycles before the friction coefficient rises 

above 0.3. It is clear that the composites have longer endurance than the 

coatlngs. However, these composltes are dlfficult and tlme consuming to 

prepare. The process involves preparing a sintered, porous metal matrix that 

is then infiltrated with molten fluorides at about 1000 "C, cooled, and flnish 

machined. If an overlay is used, the coating material Is next sprayed on from 

an aqueous slurry and then cured in an argon atmosphere at about 950 "C. 

Similar compositions can be prepared more conveniently by plasma spraying. 

Fluorlde-Metal Composite Coating 

Mixed powders of, for example, CaF2 and metal can be deposited by plasma 

arc spraying to form a composite coating on a wrought metal substrate. Excess 

coating material is applied, and the coating is then surface ground to the 

desired thickness (usually 0.010 to 0.020 cm) and a smooth surface finish. 

Two coatlngs of this type that have been successfully applied in extreme 

environments, for example, the space shuttle and the hot sectlon of small jet 

engines, are designated PSlOO and PS101. Their compositions by weight are 

PS100: 67 nichrome, 16 1/2 calcium fluoride, 16 1/2 glass 

PS101: 30 nichrome, 30 silver, 25 calcium fluoride, 15 glass 

The glass In these compositions Is a special sodium-free glass that protects the 

nichrome from oxidation. Its composition Is 58 5102, 21 BaO, 8 CaO, 13 K20. 

nal bearings. The cylindrical bores of the bearings were coated with 0.025 cm 

of PSlOO or PS101. A preoxidized, but otherwise uncoated, unlubrlcated bearing 

is included for comparison. The oxide film on the preoxidlzed bearing provided 

some protection against galllng for a tlme, but the bearing seized at 870 OC. 

Friction coefficients for the bearing lubrlcated with PSlOO were lower at all 

temperatures, and effective lubrication was achieved from about 500 to 900 "C. 

The beneficial effect of silver in reduclng low-temperature friction while only 

Figure 19 (ref. 33) gives friction coefficients In air for oscillating jour62 
moderately reducing the maximum-temperature capability of the coating system is 

tllustrated by the data for PS101. Frictlon coefficients of the order of 0.2 

were obtalned at all temperatures from room temperature to 870 OC. 

The frictlon and wear data for bearings with PSlOl lubrication, which were 

tested in moderate vacuum, cold nitrogen gas, and air, are summarized in 

table IV. The lowest bearing friction was observed in the 5~10-~-torr vacuum, 

where the friction coefficient was 0.15. Wear rates tended to decrease with 

test duration. Total diametral bearing wear was 4.5~10'~ cm after 5000 oscillatjng cycles. In cold nitrogen (-107 "C) friction coefficients were typically 

0.22 and dlametral wear after 5000 journal oscillations was 3.8~10-3 cm. 

previously discussed, the friction coeffjclent in air from room temperature to 

870 OC was approximately 0.2 over the entire temperature range. Wear rates 

a l s o w e r e uniformly low over the entire temperature spectrum. These results 

clearly demonstrate the versatility of PSlOl for lubricating plain journal 

bearings over an exceptionally wide range of temperatures and atmospheric 

conditions. 

As 

In general, the maximum useful temperature in air for a fluoride coating 

with a superalloy matrix is limited by oxidation of the alloy. Even with a 

protective glass within the composite structure, the superalloy oxidizes more 

rapidly in the composite than in the wrought metal state. The oxidation temperature limit is about 900 OC for composites with a nlckel superalloy matrlx 

and about 650 OC for cobalt matrix coatings. 

Fluoride-Oxide Composite Coatings 

Because of the limitations Imposed by oxidatton of the metal matrix coatings, completely nonmetallic coatings are of interest. It has been reported 

that plasma-sprayed coatings of Ni0 containing about 15 percent CaF2 have good 

high-temperature wear resistance (ref. 34). This coating 1s plasma sprayed 

onto seal bars for regenerators that are used in automotive gas turblne engines 

to improve thermal efficiency. The coatings must be wear resistant at high 

temperature while in sliding contact with a porous ceramic regenerator core. 

The core material is generally lithlum-aluminum silicate (LAS), magnesiumaluminum sillcate (HAS), or aluminum silicate (AS). 

Current preliminary research at Lewis Indicates that plasma-sprayed coatings based on zirconium oxide (Zr02) have attractive tribological properties. 

Figure 20 gives the friction and wear coefficients of plasma-sprayed Zr02-CaF2 

coatings with and without silver additions. In these experiments the coatlng 

was on the cylindrical surface of a rotating disk and placed in sliding contact 

with two flat nickel-base superalloy rub blocks. Both coating combinations had 

fairly high wear rates at room temperature, but wear rates were much lower for 

the ZrOz-CaFz coating at 650 OC. Silver additions were detrimental and did not 

have the beneficia1,effect of improving room-temperature friction and wear that 

they had on the metal matrix composites. Wear of the uncoated metal shoes that 

slid against the coatings was low in all cases, Indicating that the coatings 

were not particularly abrasive to the metal rub shoes. Frictlon coefficients 

were 0.4 at room temperature and 0.22+0.04 at 650 "C for the Zr02-CaF2 coating. 

Higher temperature experiments will demonstrate whether these coatings have a 

maximum-temperature advantage over the metal matrix plasma-sprayed coatings. 

63 
VACUUM-DEPOSITED COATINGS 

Vacuum-deposited coatings used in tribological applications fall into two 

main composition categories: soft lubricating coatings and very hard, wearresistant coatings. The methods of application also are in two ptlncipal 

categories: sputtering and ion plating. These techniques have been rapidly 

adopted by industry especially for aerospace applications (refs. 35 and 36). 

A very large varlety of vacuum-deposited coatings are becoming available; 

therefore only a few representative examples are discussed here. 

Soft S,olid-Lubricant Films (Sputtered) 

The most common vacuum-deposited tribological coatings are sputtered 

dlchalcogenides, especially MoS2, and ion-plated soft metals such as gold, 

silver, and lead. These coatings are often very thin, of the order of 2000 to 

5000 A in thickness. Compounds such as MoS2 are usually applied by sputtering 

because wlth proper procedures pure, essentially stoichiometric compounds can 

be deposited. In contrast, ion plating tends to dissociate chemical compounds. 

However, ion plating is an appropriate technique for depositing elemental 

metals because (1) dissociation is obviously not a problem; (2) high ion 

impact energles can be used to enhance adhesion; (3) excellent throwing power 

is achieved when coating parts with complex shapes; and (4) rapid deposition 

rates can be achieved. 

The use of sputtering to deposit MoS2 lubricating films was first reported 

in reference 37. The films were nearly stoichiometric, indicating minimal dissociation of MoS2 by the sputtering procedure used. Figure 21 shows that the 

coatlngs, which were only about 2000 A thick, had good durability in the sliding contact and low frictlon coefficients of 0.05 to 0.10 in vacuum. Unfortunately durability was much lower in air. For example, 204 size ball bearings 

with sputtered MoS2 on the cage, balls, and races easily survlved 1000 hours of 

operation In vacuum at 1750 rpm and a 138-N (31-lb) radial load, but when air 

was admitted Into the vacuum chamber, the bearings failed In less than 1 hour 

(ref. 38). Figure 22 from the same reference shows an abrupt rise in the frlction coefficient obtained with sputtered H o S 2 a s the ambient air pressure was 

increased above about 300 torr. For air applications the thicker, more conventional bonded MoS2 coatings are usually preferred; the sputter coatings are 

favored for vacuum applications. 

Soft Metal Lubricant Films (Ion Plated) 

The wear lives of ion-plated gold and vapor-deposited gold were compared In 

reference 39 with the following results. Flgure 23 shows that ion-plated gold 

had nearly twice the wear life of the vapor-deposited film. Furthermore the 

friction coefficient of ion-plated gold increased gradually, giving adequate 

indication of impending failure, while the vapor-deposited film failed abruptly 

without warning. The longer wear life of the ion-plated f i l m w a s attributed 

to its superior adherence to the substrate metal. 

The use of ion-plated lead films for lubrication of ball bearings in vacuum 

has been reported (ref. 40). The ion-plated film generated less debris and 

torque variation than vapor-deposited lead. This was again attributed by the 

authors to the superior adherence of ion-plated films. 

64 
Sputtered Hard Coatings 

Sputtered hard coating are used primarlly for wear control. The oxidation 

temperatures and hardness of some important carbides and nitrides are compared 

in table V. Coatings of all of the compounds listed are hard enough to be 

expected to have good wear resistance If adequate bonding to the substrate can 

be achieved. However, a considerable variation in oxldation reslstance exists. 

Chromium carbide, boron carbide, silicon nitride, and silicon carbide are 

oxidatively stable to at least 1000 OC; but tungsten and titanium carbides 

oxidize during long-duration exposure to air at temperatures above about 

540 "C. Tungsten carbide tends to o x i d i z e m o r e rapidly than titanium carbide 

because its oxides are volatile at high temperature and their sublimation tends 

to accelerate the oxidation. Titanium nitride is another promising hard coating material, but It too will convert t o t h e oxide above 550 "C. However, some 

TIC and TIN sputtered coatings have shown surprlslngly good resistance to oxlde 

conversion at higher temperatures than those listed In table V. 

occurs, but the rate is very low, probably because of high coating density and 

the passivating nature of the initially formed oxide film, which protects the 

coating against catastrophic oxidation. 

Oxidation 

Sputtered chromium oxide (Cr2O3) is an interesting antiwear coating. In 

start/stop tests, an optimized, sputtered coating of Cr2O3 on nickel-chromium 

foil bearings has shown outstandlng endurance over a wide temperature range 

(ref. 41). For example, the coating did not wear out after 9000 start/stop rubs 

against a journal coated wlth chromium carbtde at temperatures from room ambient 

to 650 "C (ref. 42). 

Great care is requlred in the sputtering process because reactive gases In 

the sputtering chamber can react wlth the sputtered material to alter its composition. Thls effect can be an annoyance, or it can be used to advantage In a 

process known as reactive sputtering in which controlled contaminants are 

intentionally introduced into the vacuum chamber to obtain the desired coating 

composition. It has been shown in reference 41, for example, that a controlled 

mixture of TIC and TIN can be codeposited by using a Tic target In combination 

with a partial pressure of nitrogen in the sputtering chamber. 

CONCLUDING REMARKS 

Solid lubricants for use above 300 "C were discussed. The more conventional layer lattlce solid lubricants such as MoS2 and graphite in the form of 

powders or bonded coatings are servlceabte ln:air to 350 to 400 "C under proper 

conditions and can be used to much higher temperatures for short durations as, 

for example, in s o m e m e t a l w o r k l n g processes. The high-temperature polylmldes 

have about the same upper temperature limitations as MoSZ and graphite and are 

conveniently used In the form of graphlte-fiber-reinforced, self-lubricating 

composites. For higher temperatures some soft oxides and fluorides provide 

lubrication to as high as 900 "C. They are currently used as fused coatings 

(0.001 to 0.002 cm thick) on metal substrates or as the lubricating component 

of metal matrix composites. The composites are prepared by powder metallurgy 

methods or by plasma arc spraying. For wear control sputtered hard coatings 

of some selected carbides, nitrides, and oxides are serviceable to 1000 "C if 

adequate adhesion to the substrate Is maintained at all temperatures required 

by the application. Thin, sputtered MoS2 films, typically of 2000 to 5000 A 

65 
thickness, are very effective in vacuum but of limited durablllty In reactive 

atmospheres such as air. Ion-plated films of soft metals such as gold and lead 

are effective lubrlcants partlcularly for solid-lubricated rolling contact 

bearings. 

REFERENCES 

1. Bryant, P.J., Gutshall, P.L., and Taylor, L.H., A study of mechanisms of 

graphite friction and wear, Wear, Vol. 7 (1964) 118-126. 

2. Winer, W.O., Molybdenum disulfide as a lubricant, a review of the fundamental knowledge, Wear, Vol. 10 (1965) 422-452. 

3. Sllney, H.E., High temperature solid lubricants, Mechanical engineering, 

Vol. 96, NO. 2 (1974) 18-22. 

4. Brainard, W.A., The thermal stability and frictjon of the disulfides and 

ditellurides of molybdenum and tungsten in vacuum. National Aeronautlcs 

and Space Administration, NASA TN D-5141 (1969). 

5. Matveevsky, R.M., Lazovskaya, O.V., and Popov, S.A., Temperature stability 

of molybdenum disulfide soltd lubricant coatings in vacuum. Proceedings 

2nd Int'l Conf. on Solld Lubrication 1978, ASLE SP-6, 41-44. 

6. Savage, R.H., Graphite lubrication, J. Appl. Phys., Vol. 19, No. 1 (1948) 

1-10. 

7. Peterson, M.B. and Johnson, R.L., Frict-lon studies of graphite and mixtures of g r a p h l t e w i t h several metallic oxides and salts at temperatures 

to 1000 OF, National Aeronautics and Space Administration, NACA TN-3657 

(1956). 

8. Daga, R.H., A study of high temperature lubrication by graphite in metal 

processing, Proceedings 2nd Int'l Conf. on Solid Lubrication 1978, ASLE 

SP-6, 30-37. 

9. Rudorff, W., and Rudorff, G., Structure of carbon monofluoride, Z. Anorg. 

Chem., Vol. 253 (1947) 281-296. 

thermal decomposition of polycarbon monofluorlde, Inorg. Chem., Vol. 4, 

10. Kuriakose, A.K., and Margrave, J.I.., Mass spectrometric studies of the 

NO. 11 (1965) 1639-1641. 

11. Fusaro, R.L., and Sliney, H.E., Graphite Fluorlde (CFX), - A new solld 

lubricant, ASLE Trans. Vol. 13, No. 1 (1970) 56-65. 

12. Play, D., and Godet, M., Study of lubricant properties of carbon monofluoride: (CFX),, National Aeronautics and Space Administration, NASA 

TM-75191, 1975; Translation of Etude des prqprletes lubrifiantes de 

monogluorure de graphite; (CFX),, Proceedings Colloques Int. 

C.N.R.S., No. 233 (1975) 441-450. 

66 
13. Sliney, H.E., Dynamics of solid lubrication as observed by optical microscopy, ASLE Trans., Vol . 21, No. 2 (1978) 109-11 7. 

14. Gisser, H., Petronio, M., and Shapiro, A., Graphite fluoride as a solid 

lubricant, Lubr. Eng., Vol. 28, No. 5 (1972) 161-164. 

15. McConnell, B.D., Snyder, L.E., and Strang, J.R., Analytical evaluation of 

graphite fluoride and its lubrication rerformance under heavy load, Lubr. 

Eng., Vol. 33, No. 4 (1977) 184-190. 

16. Fusaro, R.L. and Sliney, H.E., Lubricating characteristics of polyimide 

bonded graphite fluoride and polyimide thin films, ASLE Trans., Vol. 16, 

NO. 3 (1973) 189-196. 

17. Campbell, M. and Hopkins, V., Development of polyimide bonded solid 

lubricants, Lubr. Eng., Vol. 23, No. 7 (1967) 288-294. 

18. Fusaro, R.L., Molecular relaxation, molecular orientation, and the 

friction characteristics,of polyimide .films, ASLE Trans., Vol. 20, No. 1 

(1977) 1-14. 

19. Devine, M.J., and Kroll, Q.E., Aromatic polyimide composition for solid 

lubrication, Lubr. Eng., Vol. 20, No. 6 (1964) 225-230. 

20. Buckley, D.H., Friction and wear of polyimide and filled polyimide 

compositions in vacuum, National Aeronautics and Space Administration, 

NASA TN 0-3261, 1966. 

21. Lewis, R.B., Wear of polyimide resin, Lubr. Eng., Vol. 25, No. 9 (1969) 

22. Giltrow, J.P. and Lancaster, J.K., Carbon-fiber reinforced polymer as 

356-359. 

self-lubrtcating materials, I. Mech. E., 6th Tribology Group Convention, 

1968, paper 18, 149-159. 

spherical bearings to 340 OC, National Aeronautlcs and Space 

Administration, NASA TN 0-7078, 1972. 

23. Sllney, H.E. and Johnson, R.L., Graphite fiber-polyimide composites for 

24. Gardos, M.N. and McConnell, B.O.., Development of high-load, hightemperature, self-lubricating composites, Parts I - IV, ASLE Preprlnts. 

81 -LC-3A-3,4,5, and 6, 1981 . 

25. Fusaro, R.L. and Sliney, Hi€., Friction and wear behavior of graphite 

fiber reinforced polyimide composites, ASLE Trans., Vol. 21, No. 4 (1978) 

337-343. , 

26. Sliney, H.E. and Jacobson, T.P., Performance of graphite-fiber-reinforced 

polyimlde composites in self-allgning plaln bearings to 315 OC., Lubr. 

Eng., Vol. 31, No. 12 (1975) 609-613. 

27. Sllney, H.E., Some load limits and lubricating properties of plain 

spherical bearings with molded graphite-fiber-reinforced polyimide liners 

to 320 "C, Lubr. Eng. Vol. 35, No. 9 (1979) 497-502. 

67 
28. Williams, F.J., Composite airframe journal bearings,.Rockwell 

International, NA-80-648, National Aeronautics and Space Administration, 

NASA CR-165249, 1981 

29. Sllney, H.E., Effect of sliding velocity on friction properties and 

endurance life of bonded lead monoxide coatings at temperatures up to 

1250 OF, National Advisory Committee for Aeronautics, NACA RM E 58 611, 

1958. 

30. Sliney, H.E., Strom, T.N., and Allen, G.P., Fluoride solid lubricants for 

extreme temperatures and corrosive environments, ASLE Trans., Vol. 8, 

NO. 4 (1965). 307-322. ' 

31. Sllney, H.E., Rare earth fluorides and oxides - An exploratory study of 

their use as solid lubricants at temperatures to 1800 OF (1000 "C), 

National Aeronautics and Space Administration, NASA TM 0-5301, 1969. 

32. Sliney, H.E., Self-lubricatlng composites of porous nickel and 

nickel-chromium alloy impregnated with barium fluoride - calcium fluoride 

eutectic, ASLE Trans., Vol. 9, No. 4 (1966) 336-347. 

33. Sliney, H.E., Wide temperature spectrum self-lubricating coatings prepared 

by plasma spraying, thin solid films, Vol. 64 (1979) 211-217. 

34. Moore, G.D. and Rltter, J.E., The friction and wear characteristics of 

plasma-sprayed NIO-CaF2 in rubbing contact with a ceramic matrix, Lubr. 

Eng., Vol. 30, No. 12 (1974) 596-604. 

35. Christy, R.K. and Barnett, G.C., Sputtered MoS2 lubrication system for 

spacecraft gimbal bearings, Lubr. Eng., Vol. 34, No. 8 (1978) 437-443. 

36. Gardos, M.N., Quallty control of sputtered MoS2 films, Lubr. Eng., 

VOI. 32, NO. 9 (1976) 463-480. 

37. Spalvins, T. and Przybyszewski, J. S., Deposition of sputtered molybdenum 

disulfide films and friction characteristics of such films In vacuum, 

National Aeronautics and Space Adminlstration, NASA TN 0-4269, 1967. 

38. Spalvins, T., Bearing endurance tests.in vacuum for sputtered molybdenum 

disulfide films, National Aeronautics and Space Administration, NASA 

TM X-3193, 1975. 

39. Spalvins, T., Coatings for wear and lubrication, Proceedings 3rd Int'l. 

Conf. on Metallurgical Coatings, Elsevier, 1979. 

40. Todd, M.J. and Bentall, R.H., Lead film lubrication In vacuum, ASLE SP-6 

( 1978) 148-1 57. 

41. Bhushan, 6.. Development of surface coatings for air lubricated bearings 

to 650 "C, ASLE Trans., Vol. 23, No. 2 (1980) 185-196. 

42. Brainard, W.A. and Wheeler, D.R., Use of nitrogen-argon plasma to improve 

adherence of sputtered titanium carbide coatings on steel, J. Vac. Sci. 

Technol., Vol. 16, No. 1 (1979) 31-36. 

68 
TABLE 11. - TYPICAL GRAPHITE FIBER PROPERTIES 

lProperty or Type "L" Type "H" 

Character i st i c 

Tensile strength 

Elastic Modulus 5.0~10~ lb/in2 3x1O1O N/m2 5.7~10~ lb/in2 3.9~10~~ N/m2

Length 0.25 in. 6.4~10'~ m 0.25 in. 6.4~10'~ m 

Diameter 3.3xW4 in. 8.4~10~~ m 2.6~10'~ in. 6.6~10'~ m 

Specific gravity 1.4 1.4 1.4 1.4 

English Units SI Units English Units SI Units 

9.0~10 4 lb/in2 6.2~10~ N/m2 2.8~10~ lb/in2 2.0~10~ N/m2 

TABLE 1. - RESULTS OF THERMAL STABILITY AN0 FRICTIONAL 

EXPERIMENTS IN VACUUM OF 10-9 to 10-6 TORR 

Compound Probable Dissociation Maximum 

onset of products temperature 

thermal e first detected at which 

di ssoci at i on by mass burnished 

as detected spectpnetry, films provided 

by C effective 

C lubrication, 

C 

MoS2 930 1090 650 

870 1040 730 

760 980 760 

700 930 760 

700 700 540 

ws2 

wse2 

ne2 700 700 (a) 

MoSe2 

MoTe2 

aFriction coefficient greater than 0.2 at a1 1 temperatures. 

TABLE 111. - COMPARATIVE WEAR LIFE OF FLUORIDE COMPOSITES AND 

COATINGS IN AIR AND HYDROGEN 

Specimen Cycles at which friction coefficient 

temperature increased to O.3Oa 

Air Hydrogen 

Composites Coatings Composites Coatings 

25' (b) (c) dl 560 000 (c 1 

260 2 750 000 115 000 dl 499 000 (c 1 

650 1 370 000 (c) dl 370 000 (c) 

540 1 105 000 389 000 dl 610 000 275 000 

816 850 000 (C) 570 000 (c) 

aBased on single runs. 

bLow wear rate but friction coefficient of 0.30 to 0.35. 

CNo test. 

dExperiments terminated before failure. (Friction coefficient did net increase to 0.3 during number of cycles 

i ndi ca ted . ) 

69 
TABLE IV. - PERFORMANCE SUMMARY FOR OSCILLAnNG PLAIN 

SLIDING BEARINGS SELF-LUBRICATED WITH A PLASMASPRAYED COATING IN VARIOUS ATMOSPHERES 

[PSlOl Coating: 30 Ag, 30 NiCr. 25 CUP, 15 glass: 0.025 cm 

(0.010 h.) thfck; 3.5~107 N/m2 (SO00 psi) unit load, tlSO 

oscillation at 1 bertz.] 

temperature atmosphere I Ambient 

OC I OF I ~~ 

Room 

-107 

Room 

540 

650 

870 

Room 

-160 

Room 

1000 

1200 

1600 

Vacuum 

Nitrogen 

Air 

760 torr 

s10-2 torr 

Typical 

ftfctlon 

:oefficlent 

0.15 

.22 

.24 

.19 

.21 

.23 

Increase in radial clearance ~ 

cm x103 (milliinches) 

After 100 cycles 

1.3 (0.5) 

0.3 (0.1) 

.5 (0.2) 

.5 (0.2) 

.3 (0.1) 

.3 (0.1) 

After 5000 cycles 

4.5 (1. e) 

3.8 (1.5) 

7.0 (2.8) 

2.5 (1.0) 

2.5 (1.0) 

6.0 (2.4) 

TABLE V. - BULK PROPERTIES OF SOME HARD COAT MATERIALS' 

Material Microhardness, Oxidation iemperatureb, 

k glmd C 

B4C 

Tic 

Sic 

Cr3C4 

wc 

Si 3N4 

Ti N 

C r203 

4200 

3200 

2900 

2650 

2050 

2000 

1950 

'1800 

1090 

540 

1650 

1370 

540 

1400 

540 

- 

aData from: Engineering Properties of Ceramic Materials, 

bTemperature for appreciable detrimental oxidation (passiCEstimated conversion from published Moh hardness of 9. 

Battelle Memorial Institute, Published by American 

Ceramic Society, Columbus, Ohio, 1966. 

vating oxide films form at lower temperatures). 

70 
10 000 

il 

1000 

(a) I I I I 

100 

- 1 WEEK r\ 

- 1 DAY 

- 1 HOUR 

AIR FLOW 

- 333 

A 2000 

- 1 DAY 

- 1 HOUR 

AIR FLOW 

- 333 

A 2000 

.4 

.2 

0 

RATS cclmin 

TEMPERATURE, % 

(a) Oxidation characteristics d MoS2 at two air flow rater 

Ib) CompmHv, oxidation d WS2 and M0S2 air flow rate, 

Flyre L - Oxidation kinetics d MoS2 and WS2 averalp 

113 Llmin. 

portlcle size 1 p, compact density; 50 % 

I I 

.6 r 

AIR ATMOSPHERE MELTING 

POINT 

WWO, I 

I 

I ,-MELTINGPOINT I 

I 

re OF MOO^ I 

1 II 

0 

TEMPERATURE, % 

Figure 2. - Friction characteristics of hS2and WS2 in argon and 

in air. 

71 
I 

z 

P 

u_ 

m 

8 

Y 

0 INCREASING TEMPERATURE [I 0 DECREASING TEMPERATURE 

0 CADMIUM OXIDE ALONE 

TEMPERATURE, ‘% 

(a) Graphite alone, 

(b) Cadmium oxide - graphite mixture. 

Figure 3. - Effecl d wide adjuvant on lubrication with 

graphite. 

o FLUORINE 

CARBON 

T 

1 

(a) Graphite. (b) Graphite fluoride. 

Figure 4: - Structure Ol graphite and proposed structure of 

graphite fluoride illustrating the expansion d the carbon 

layer planes due to the intercalation of fluorine atoms. 

72 
POWDER 

1 

,? UNLUBRICATED METAL - -6 I-' 

z 

Y, 

DECOMPOSITION 0 

ut 

I I 

0 loo 200 300 400 500 600 

TEMPERATURE, OC 

Figure 5. - Effect of temperature on wear life and 

friction coefficient of graphite fluoride 

IICF1. 

burnishe! on sandblasted 440-C stainlesssteel disks.. Riders, 440-C stainless steel; lin - 

ear sliding speed, 1.6 mls; load, 500 $ atmosDhsre. dry air (moisture content, 20 wm). 

) and molybdenum disulfide powders 

1.. 

br, -2, e.: 

(a) FIRST PARTICLES ENTERING 

CONTACT. 

(c) PROGRESSIVE FILM 

FORMATION. 

, . 

(b) PROGRESSIVE FllM 

FORMATION. 

(d) COMPLnE FILM FORMATION 

AFlER ONLY IO mm OF SLIDING. 

Figure 6. - Behavior of graphite fluoride in initially unlubricated contact. 

Load, 13.2 N (3 Ib); original magnification, X150. 

73 
la) Sic PARTICLE ENTERING 

CONTACT. 

- 

* 

*: *. - 

.* r 

lc) PARTICLE AT CONTACT 

EXIT. 

(bl PARTICLE PART WAY 

THROUGH CONTACT. 

(dl CONTACT AFlER 10 DISK 

REVOLUTIONS. 

Figure 7. - Abraslve action of silicon carbide particles. Load, 13.2 N 

(3 Ib): original magnification, X1.50. 

0 DRY ARGON (10 ppm. MOISTURE) 

0 DRY AIR (20ppm) 

A MOIST AIR (10 OOO ppm) 

10 OOO 

1OOO 

100 

GON 

10 

IIN AIR 1 1 

.1 

0 100 200 300 ooo 500 m 

TEMPERATURE, % 

Figure 8. -Friction coefficient and wear life as a 

function of temperature for thin films d polyimi& run in atmospheres d dry argon, dry air, 

and moist air. Load, 1 kg; velacity. 3 mlsec. 

i 

74 ORIGINAL' PAGE 

BLACK AND WHITE PHOTOGRAPH 
10 m c o POLYIMIDE' (PI) 

A PI-BONDED (CF 1) 

0 PI -BONOEO Mo'k 

h UNLUBRICATED ' 

I I I I 1 I I 

100 200 300 400 m 600 

TEMPERATURE, '% 

0 

Figure 9. -Friction ccefflcient and wear life as a function 

d temperature.for three solid lubricant films run in dry 

air (moisture content, 20 ppml. Load, 1 k$ velocity, 

3 mlssc 

n 

(b) 

(a) Condensation-type of polyimide polymer (type "C"). 

(b) Addition-type of polyimide polymer (type "A"). 

Figure la -Two major types of polyimide structures. 

75 
t 

0 

U 

10 10 -300 loo 

NUMBER OF REVOLUTIONS. kc 

(a) Friction coefficients. 

MPE OF 

COMPOS ilE 

1x10-1 

I 

cm 

250 C 3oooC 

(b) Surface profiles of wear tracks after 300 OOO cycles of sliding. 

Figure 11. - Friction and wear of graphite fiber reinforced polymide composites 

in moist air atmospheres (10 OOO ppm H20). 

LGRAPHllE FIBER REINFORCED 

POLYIMIE BALL 

T GRAPHITE FIBER REINFORCED ' POLYIMIDE LIFER 

-44X-HT BALL 

MC-HT OUTER RING 

Figure 12. -Test bearings employing graphite fiber reinforced 

polyimide. 

76 
A BALL FRACTURE 

0 CONVENTIONAL PTFE-LINED BEARING 

(MS 21233- 10) - GRAPHITE- POLYIMI DE COMPOSITE BALLS 

GRAPHITE 

w 

E & .2 

A Y & .2 15- 

IVU 

A FAILURE 

ERMTRUDES I ~ 

0 100 200 300 400 

BEARING TEMPERATURE, (5 

figure 13. - Summary d friction d spherical bearing with 

polyimide - graphite-fiber composites d various fiber 

con nts. Stellite 68 journal; radial unit load, 15x10' 

at 1 Ht. *lp. 

Nlm 'I! (5NI psi); journal oscillation in cylindrical bore 

-I 0 MOLDED LINER 

0 MOLDED BALL 

.2c T .It p" I I $ I ,$, , 

0 200 300 400 

TEMPERATURE, bc 

Figure 14. - Fridlon-temperature characteristics of 

two bearing designs lubricated with GRPI composite. 

OL 1 m 400 

0 100 200 

TEMPERATURE. OC 

PLAIN CYLINDRICAL 

PTR LINER 

Figure 15. - Dynamic unit load capacities of three bearing designs 

self lubricated with GFRPI composite. 

77 
10-10 

TEMPERATURE, % 

Figun 16. - Effect of sliding velocity and temperature 

on friction properties of bonded PbO-SiOt. 0.003 cm 

thick; load, 10Wg: 

DISK SLIDING 

VELOCITY, 

mlsec 

-UNCOATED 2 

0 COATED 2 

0 COAED 10 - 

0 250 m 750 loo0 

TEMPERATURE, % 

Figure 17. - Lubricating properties of 0.003 cm 

thick fused fluoride coating composition in 

air. Load, 500 5 

78 
I- 

.4 

.2 

COMPOSITES 

0 COATED 

0 UNCOATED 

0 UNLUBRICATED 

0 

RIDER WEAR 8 

I I I I 

0 250 m 750 lW0 

TEMPERATURE. '% 

figure 18, - Frldlon and wear of fluorb-Inconel 

composite disks and cart Inconel riders in air. 

(35 vol % Bd2-Cd eutectlc, 65 vol % sintered 

Inconel: m-g I&, 10 mlsec. 

A PREOXIDIZED UNLUBRICATED RE& 41 

BEARING 

(NICHROMECALCIUM FLUORIDE GLASS) 

0 BEARING BORE COATED WITH PSlW 

0 BEARING BORE COATED WITH PSlOl 

.8 wsim WITH SILVER ADDED) 

lI I I I I 

0 200 400 600 800 loo0 

BEARING TEMPERATURE. OC 

Figure 19. - Bearing friction. 

79 
9 80 ZIRCONIA: 20 CALCIUM 

FLUORIDE 

-80 ZIRCONIA: 10 SILVER: 

25 10 CaF2 

25 C 

Flgure M -Wear and friction of plasma - sprayad coatings d zlrconia and calcium 

fluorIC, with and without silver. Double 

rub shoe tests with 22.7 kg per Inconel 

shoe agalnst coated disk at 0.3 mlsec 

(1% rpm). 

L3 

E L2 

Iu I I I I 

.1 - 

L I I 

0 40 80 120 160 200 240 280 320 

TIME, min tb) 

la) Specimen, niobium. 

(b) Specimen, nickel-chrome alloy. 

Flgure 21. - Averag friction coafficlents d nioblum sliding 

on two different specimens coated with sputtered m lybdenum disulfide in vacuum I 0 torr or 1.33~10- 8 N/m21. Load, 250 g; speed, 2.5ox10- 1 mlsec; -11 ambient temperature. 

80 
0 INCREASING PRESSURE 

0 DECREASING PRESSURE .14 a 

U 

PRESSURE, torr 

Figure 22. -Effect d pressure on coefficient of slldlng 

friction for sputtered MOST Load, 250 9; speed, 

Q 11 mlm substratelrider. NiINi; room temperature. 

COmlCiEM OF FRICTION 

FOR BARE (Ni-Cr) 

SUB STRATE 

VAPOR 

DEPOSITION ION 

V PLATING 

.2 

0 20 40 60-80 100 

u 

rnin 

2200 6600 7000 

CYCLES 

Figure 23. - Coefficient d friction d niobium 

sliding on (Ni-Cr) alloy with gold deposited 

by vapor deposition. and Ion plating about 

2000 A thick (load, 250 y peed, 1.52 mlmln; 

ambient temperature 10-7ftorr). 

81 
Mechanical Components 

Bearings 

Gears 

Traction Drives 

a3 
DESIGN AND LUBRICATION OF HIGH-SPEED ROLLING-ELEMENT BEARINGS* 

Erwin V. Zaretsky 

National Aeronautics and Space Admlnlstration 

Lewis Research Center 

Cleveland, Ohio 44135 

The speed capability of rolling-element bearings has increased from speeds 

of less than two million DN to speeds of three million ON. The life and rellability of these bearings have also increased where they are equal to, or 

greater than, those of bearings with.1imited speed capability. However, highspeed bearings are not "off-the-shelf" bearlngs which can be readily ordered 

from a manufacturer's catalog. Design parameters must be carefully chosen and 

optimized based upon sophisticated bearing computer programs. Material and 

lubricant selection must be integrated Into the bearing design. Bearing 

thermal management must be implemented through proper lubrication and cooling. 

Parameters which can be used to design, specify, and lubricate hlgh-speed 

bearings are presented and dtscussed. 

INTRODUCTION 

In the early years of the aircraft turbojet engine, engine life and 

rellability were generally llmited by rolllng-element bearlng technology. 

Bearing life in a jet engine environment was limited to approximately 300 hr 

MTBR (mean time between removal). Research and development activities by the 

major U.S. engine manufacturers, bearing companies and government laboratories 

over the past three decades have resulted in extending bearing lives in engine 

operation to approximately 30 000 hr MTBR. At the same time bearlng speed 

capability has also been vastly increased (fig. 1). Bearing speeds of 3 milllon ON and lives 100 times catalog rating have been demonstrated for speclally 

designed rolltng-element bearings using material, lubrication and manufacturing 

technologies which are now commercially available. 

demand the same design and lubrication considerations as those found ln 

advanced airbreathing aircraft engines. Besldes the effect of higher stress 

at the contact of the rolling element and the outer race due to centrifugal 

force at higher speeds, heat generation and power loss within the bearings are 

major problems. Thermal analysls of the Internal geometry of the bearlng with 

regard to fits and clearances becomes lmportant. Shaft and houslng tolerances 

on the fit and expansion of the bearing rings are a consideratlon. Cooling of 

the bearing through lubrication must be efficiently performed. Materlal and 

lubricant type must be selected to both assume compatibility wqth each other 

and with the higher temperatures found in hlgh-speed bearlngs. Current lubricants are limited to bearing temperatures of 425 OF and materials to temperatures above 600 OF. 

Bearings used in commercial turbomachinery operating at higher speeds 

r 

In designing a bearing for these higher speeds a balance must be malntalned among attalning the required life, incurring surface damage by skidding 

*Prepared 'or the Ortgtna 1 Equlpment Manufar turing Design Conference, Phi ladelphia, Pennsylvania, September 9-11, 1985 (NASA TM-87107). 

85 

PRECEDING PAGE BLANK NOT FILMED 
of the rolling elements, inner-ring fracture due to resulting hoop stresses 

and fretting problems due to improper fits at operating conditions. 

deflections from shaft loading and rotor dynamics can exceed the bearing design 

limitations. When a tolerance of 100 millionth (0.0001) of an inch Is critical 

to bearing operation and performance these deflections become an Important 

design consideration. 

Shaft 

The design of a high-speed rolling-element bearing is a rather sophisticated process. Contrary to the implication of many engineering design texts, 

it is not possible to select a high-speed bearlng from a manufacturer's catalog 

cbnsidering only the external dimenslons of the bearing and the applied load 

and speed. These bearings must be designed for the high-speed application and 

conversely, the application must be designed for the bearings. Sophlstlcated 

bearing computer analysis must be employed for this purpose. It becomes the 

objective of this paper to review the design and lubrication methods currently 

used in high-speed, rolling-element bearing design and lubrication as well as 

material and lubricant selection. While there may be other design methods and 

analysis employed for this purpose, the methods presented have been established 

in laboratory experiments and field service to achieve the desired performance 

results. 

COMPUTER ANALYSIS 

There are several comprehenslve computer programs that are capable of 

predicting rolling-element bearing operatlng and performance characteristlcs. 

These programs generally accept input data: bearing internal geometry (such . as sizes, clearance, and contact angles), bearing material and lubrlcant properties, and bearing operating conditions (load, speed, and ambient temperature). The programs then solve several sets of equations that charactertze 

rolling-element bearings. The output produced typically consists of rollingelement loads and Hertz stresses, operating contact angles, component speeds, 

heat generation, local temperatures, bearing fatigue life, and power loss. 

The critical assumptions currently necessary In the use of these programs are 

the form of the lubricant traction model and the lubricant volume percent (the 

assumed volume percent of the bearing cavity occupied by the lubricant). Two 

of these programs titled "Shaberth'' and "Cybean" are available from COSMIC, 

112 Barrow Hall, University of Georgia, Athens, GA 30602, or SKF Industries, 

Inc., 1100 First Ave., Klng of Prussia, PA 19406. 

The Shaberth program simulates the thermal-mechanical performance of a 

flexible shaft supported by as many as five rolling-element bearings comprising combinations of ball, cylindrical and tapered roller bearlngs. Cybean 

analyses a single cylindrical roller bearing. Both programs are capable of 

calculating the thermal and kinematic performance of high-speed bearings, and 

Cybean includes a roller skew predlctlon for misaligned condltlons. Flgure 2 

illustrates the modal system for thermal routines In Cybean. The correlatlon 

between predicted and experimental results with these two programs Is very 

good. However, the proper use of these programs requires an engineer knowledgeable in bearing technology. 

MATERIAL SELECTION 

A commonly accepted minimum hardness for rolling-element bearing components at operating temperature Is Rockwell C58. At hardnesses below this 

86 
value brinelling of the bearing races can occur at stress levels normally 

experienced. Bearing life is also a functlon of material hardness. Since 

hardness decreases with temperature, conventional bearing materials , such as 

AISI 52100 can be used only to temperatures of about 250 OF. The M-series 

steels such as AISI H-50 can retain a minimum hardness of Rockwell CS8 to 

approximately 700 OF. Because of Its life potential and hardness retention, 

the AISI M-50 material is the choice material for high-speed ball and roller 

bearings. The nominal room temperature material hardness of the rolling elements and the races should be Rockwell 63. The hardness difference (AH) of 

the rolling elements minus the races should be zero. This material in its 

through hardened state can be used to speeds of approximately 2.3 mi.llion DN. 

At higher speeds with the occurrence of a spa11 In the inner race, hoop 

stresses due to centrifugal force and press fits can cause the inner ring to 

fracture causing catastrophic failure of the bearing and probable secondary 

failure of the rotor system. Clearly this failure mode Is unacceptable. As a 

result, bearing speeds using through hardened steels must be limited speeds of 

less than 2.3 million ON. 

A solution.to this problem is the use of carburized or surface hardened 

steels wh'erein the material core remains in a soft ductile state and fracture 

resistant. E.N. Bamberger at the General Electric Company, devised a modified 

AISI M-50 material which can be carburized. This material is designated M-50 

Nil. With the M-50 Nil material, bearing speeds of 3 million DN can be 

achieved without fear of Inner-ring fracture. 

percent or less. This should minlmize,dimensional growth because of the 

retained austenite transforming into martensite. 

Retained austenite content of the AISI M-50 material should be kept at 3 

Roll ing-element bearing re1 iabi 11 ty and load capabi 1 i ty increases significantly when nonmetallic inclusions, entrapped gases, and trace elements are 

eliminated or reduced. Improv,ements in steel making processing, namely double 

melting in a vacuum, can achieve this. Vacuum-induction melted, vacuum-arc 

remelted (VIM-VAR) AISI M-50 steel achieves the aforesaid results wjth significant improvement in bearlng life and reliability over single-vacuum melted 

(VIM) steel. VIH-VAR AISI M-50 and VIM-VAR M-50 NI1 are commerclally available 

and should be specified for high-speed bearing application. 

For angular-contact ball bearings, the bearing races should be forged to 

assure parallel grain flow in the ball-race contact. Research has shown that 

improvement in life of rolling-element bearings occurs with parallel grain 

flow. For cylindrical roller bearings and deep groove ball bearings, the 

preferential grain flow already exists wlthout the need for forging. 

In recent years the use of ceramlc rolling elements has been proposed for 

high-speed bearing applications because of their lighter weight resulting In 

lower centrifugal force at the outer-race contact. What has not been considered is that the modulus of elasticity (Young's modulus) and/or Poisson's 

ratIo of these materials are higher than steel such that, even consldering the 

reduction In centrifugal force, the contact stresses at the Inner and outer 

races are greater than with steel rolling elements. This would result in lower 

lives than with an all steel bearing. Unless there.are conditions where the 

environment is noncompatible with steel, ceramic materials are not recommended 

for high-speed bearing application. 

87 
BEARING DESIGN 

Angular-Contact Ball Bearings - Angular-contact ball bearings can be 

designed and operated for speeds of 3 million ON with acceptable llfe and 

reliabillty. Based upon experience, a nominal contact angle of 24" and race 

comformities of 54 and 52 percent at the inner and outer races, respectlvely, 

can be used as starting parameters in the design. These parameters can be 

varied in the design processes to optlmlze fatigue life and minimize power loss 

and temperature differences between the inner and outer races. 

contact angle I3 will change with load and speed as shown in fig. 3. As speed 

is increased, the contact angle at the outer race will decrease while the contact angle at the inner race will increase. The bearing designer or user must 

determine whether at full operating speed and load the ball will ride on the 

shoulder of the inner race. Should thls occur, bearing life can be significantly shortened. The differences in temperature between the inner and outer 

races wlll affect the internal bearing clearances and, thus, bearing performance and llfe. Should the inner-race temperature increase at a greater rate 

than that of the outer race, the bearing may lock up. 

The nominal 

For high-speed applications wlth bore slzes greater than 75 mn, the cage 

design should be a one-plece Inner-land riding type. Experlence has lndlcated 

that the cage should be made out of an iron base alloy (AMS 6415) heat treated 

to a Rockwell C hardness range of 28 to 35 and havlng a 0.005-cm (0.002 in) 

maximum thickness of silver plate (AMS 2410). 

within 3 gm-cm (0.042 oz-in). For bore sizes less than 75 mn, single-land, 

outer-race riding cages may be the preferred design. 

The cage should be balanced 

The tolerance grade of the bearings should be ABEC-5 or better. 

the ABEC specification does not consider surface finish or waviness. 

of the raceway should be 100 pin or less. 

be 2 pin or better and l'pin for the balls. 

1 pin can be obtained for the raceways. However, not all precision bearing 

manufacturers can attain thls finish. ' 

However, 

Waviness 

Surface finishes of the races should 

With honing, surface finishes of 

Roller Bearings - High-speed roller bearings present perhaps a greater 

design challenge than ball bearings. 

3 million ON. For these bearings at high speed, surface distress caused by 

roller skidding or skewing is the cause of failure. Skldding occurs when the 

radial load on the bearing is Inadequate to develop enough tractive force 

between the raceways and roller. With insufflclent tractive force, true rolling cannot be maintained, and roller sliding or skidding results. Where there 

Is a sufficient lubricant film separating the surfaces no damage to the bearlng 

would be anticipated. However, where the film is marginal, damage can occur. 

These bearings can be run to speeds of 

Skewing of a roller is erratic gyration on a axis not parallel to the axis 

of the bearlng. Skewing results In loading of the roller ends producing excessive end wear, wear of the race flanges or guide channels, and wear of the cage 

or separator pockets. Should this condition continue for any length of time, 

cage fallure will occur or the roller wlll wear sufficiently to turn 90" in 

the directlon of rolling and lock up the bearlng. Hence, the bearing must be 

designed to prevent skewing and minimize skidding. 

The primary method of eliminating s-kidding is by lncreaslng the contact 

stress on the indlvidual rollers. This can be done by reduclng both the size 

88 
. 

and number of rollers in a bearing. The amount of reduction will be limited 

by the desired design life of the application. 

Another method for increasing the contact stress Is shown in fig. 4. In 

most roller bearings, the maximum number of rollers under load is 20 percent 

as illustrated In fig. 4(a). Deflecting the bearing outer race by adding a 

radial preload at two points 90' to the applied load the number of rollers 

under load can be increased to approximately 60 percent. This is illustrated 

in fig. 4(b). The design and manufacture .of an elliptical outer raceway can 

accomplish this result. 

There have been suggestions to use preloaded hollow rollers to produce a 

load on all rollers within the bearing. While at first blush this suggestion 

has merit, experiments have shown that these hollow rollers can fall by flexture fatigue resulting in catastrophic bearing failure. 

For most roller bearings, the roller of the bearing is crowned to prevent 

edge loading of the roller ends and thus premature failure. Typical features 

of a crowned roller for a high-speed bearing are shown in fig. 5. The length 

of the roller should not exceed its diameter. The corner radius runout with 

the diameter of the roller should not exceed 0.001 in total. Concentricity of 

the crown profile to the diameter should be within 25 millionth of an inch. 

Roller ends should be s q u a r e w i t h i n 0.0001 in. Surface finishes on the rollers 

are usually 2 to 5 pin and on the raceway 4 to 6 pin. 

modate a l a r g e m i s a l i g n m e n t and prevent skewing. This misalignment Is caused 

by improper assembly or shaft bending under load. However, Increasing crown 

drop beyond 0.007 in will not greatly affect skewing. 

Referring to fig. 6, a relatively large crown drop Is necessary to accomUnbalance of the roller will cause roller-gyration skewing. Where the 

roller Is not dynamically balanced then roller-gyration skewing may be caused 

by (a) excessive corner radius runout with the roller diameter; (b) a crown 

profile skewed to axis of the roller and (c) lack of squareness of the ends of 

the roller to the diameter. 

Cage design considerations should be similar to those of the ball bearings. Clearance of the roller in the cage pocket should be approximately 

5 percent of the roller diameter to permit the roller to move within the pocket 

without excessive wear. Speed variations cause this roller movement. 

Channel-to-roller clearances must be small to maintaln good guidance of 

the roller. Total clearance should be approximately 0.0010 to 0.0015 in. 

This clearance is sufficient to allow for.lubrication of the roller end - 

flange contact and provides for adequate guidance. 

Frederick T. Schuller of the NASA Lewis Research Center has developed a 

three piece inner-race roller bearing (fig. 7). This concept allows the 

flanges and races to be manufactured from different materials where required 

and to be machined with super finishes on the raceway and the roller-flange 

contacts on the order of 1 to 2 pin RMS. The concept eliminates stress concentrations at lubricant holes in the raceway surfaces at hlgher speeds found 

in conventional designs. It also allows for modlfled configurations of the 

flange and roller (fig. 8). This bearing deslgn has been successfully run 

89 
to speeds of 3 million ON. While the bearing is not comnercially available, 

it can be specially ordered from precision bearing manufacturers. 

Tapered-Roller Bearlng - Tapered-roller bearings have been restricted to 

lower speed applications than have been ball and cylindrical roller bearings. 

The speed limitation is primarily due to the cone-rib/roller-end contact which 

requires very careful lubrication and cooling consideration at higher speeds. 

The speed of tapered roller bearings is limited to approximately 0.5 million 

DN (a cone-rib tangential velocity of approximately 700 ft/min) unless special 

attention Is given to lubricating and designing this cone-rlb/roller-end contact. At higher speeds centrifugal effects starve this critical contact of 

lubricant. 

A computer optimized design of a tapered roller bearing capable of speeds 

to 2.4 million ON is given in table 1. Tapered roller bearings must use case 

carburized steels. For higher speed applications materials such as CBS-1000 M 

and CBS 600 should be considered. To date, M-50 Nil has not been used for 

tapered roller bearings. 

There is a caveat for using tapered-roller bearings at higher speeds. 

Should there be lubricant starvation to the bearing, the bearing will fail in 

an extremely short time period. This time is usually less than the time to 

shut down the system without secondary damage. The problem Is currently being 

worked on. However, no state-of-the-art solution to a commercial user is 

available. 

LUBRICATION METHODS 

Jet Lubrication - Where speeds are too high for grease or simple splash 

lubrication, jet lubrication is used to both lubricate and control bearing 

temperatures by removing generated heat. 

number of nozzles, jet velocity, lubricant flow rates, and removal of lubricant 

from the bearing and immediate vicinity are all very important for satisfactory 

operation. 

In jet lubricatlon the placement and 

The placement of jets should take advantage of any. natural pumping ability 

of the bearings. This is illustrated in fig. 9 for a ball bearing with 

relieved races and for a tapered-roller bearing. Centrifugal forces aid in 

moving the oil through the bearing to cool and lubricate the elements. 

Directing jets in the radial gaps between the races and the cage is beneficial. The design of the cage and the lubrication of its surfaces sllding on 

the shoulders of the races greatly affects the high-speed performance. The 

cage has been typically the first element to fail in a hlgh-speed bearing wtth 

improper lubrication. 

It has been shown that with proper bearing cage design, nozzle placement, 

jet velocities, and adequate scavenglng of the lubricant, jet lubrication can 

be successfully used for small-bore ball bearings at speeds to 3 million ON. 

For large ball bearings, speeds to 2.5 million ON are attalnable. For large 

tapered-roller bearings, jet lubrication was successfully demonstrated to 1.8 

million-ON, although a high lubricant flow rate of 0.0151 m3/min (4 gpm) and 

a relatively low oil-inlet temperature of 170 OF were required. 

90 
Under-Race Lubrication - A more effective and efficient means of lubricating rolling-element bearings is under-race lubrlcation. Conventional jet 

lubrlcatlon fails to adequately cool and lubricate the inner-race contact as 

the lubricant is thrown Centrifugally outward. Unfortunately, lncreased flow 

rates add to heat generated from oil churning. An under-race lubrication system used In turbofan englnes for ball and cylindrical roller bearings is shown 

In flg. 10. Lubricant 1s directed under the inner race and centrifugally 

forced out through a plurality of holes in the race to cool and lubricate the 

bearlng. 

only as shown in fig. lO(a). Lubricant supply holes are usually provlded for 

the cage-land and the roller-flange contacts (fig. 11). 

Some lubricant may pass completely under the bearing for cooling 

Proper "Bearing Thermal Management" Is a requirement for successful highspeed bearing operation. Thls can be further achieved by outer-race coollng 

with under-race lubrication (fig. 12). By control of lubricant flow to the 

outer- and inner-races internal clearances of the bearing are maintained and 

controlled. over a range of loads and operating speeds to 3 million DN 

(fig. 12(a)). For the tapered-roller bearing (fig. lZ(b)), by supplylng 

lubricant t o t h e cone-rlb/cage contact as well as the cone-rib/roller-end, 

contact speeds of 2.4 million DN can be reached. 

LUBRICANT SELECTION 

The criteria for a liquid lubricant to function in a rolllng-element 

bearlng are that (a) It be thermally and oxidatively stable at the maximum 

bearing operating temperature, and (b) it form an elastohydrodynamlc (EHD) 

film between the rolling surfaces. The EHD film, which is generally dependent 

on lubrlcant base stock and viscosity, is 5 to 100 pin thick at high temperatures. When a sufficiently thick EHD film is present, rolling-element bearings 

will not usually fail from surface distress. Instead, they fail from rolllngelement fatigue which usually manifests itself, in the early stages, as a 

shallow spa11 with a diameter about the same as the contact wldth. 

the 

f in1 

thic 

( a? 

A requirement for long-term high-temperature bearing operation is that 

EHD film thlckness, h, dlvlded by the composlte surface roughness, 

t u$)1/2, equal 1-1/2 or greater, where a1 and 02 are the surface 

shes of the raceway and rolling elements, respectlvely. The EHD film 

kness is a function of several lubrlcant and bearlng operating varlables. 

However, as a general rule, the minimum viscoslty required of a lubricant is 1 

cSt at operating temperature. Thls same research indicated that the ester 

based lubricants (table 2) meeting the MIL-L-23699 specification could provide 

the necessary lubrication requlrements to 425 OF in an alr environment. While 

other base stock lubrlcants could give satisfactory operation to 600 OF, they 

were precluded from further conslderation because of their cost and/or commercial availabtllty. Further, at temperatures above approximately 450 OF a low 

oxygen environment would be required to minimize lubrlcant oxidation for most 

lubricant types. 

The speed capability of 

of less than 2 mllllon ON to 

of these bearings have also 

CONCLUSION 

rolling-element bearlngs has lncreased from speeds 

speeds of 3 mlllion DN. The life and rellablllty 

ncreased where they are equal to, or greater than, 

91 
those of bearings with limited speed capability. 

are not "off-the-shelf" bearings which can be readily ordered from a manufacturer's catalog. 

based upon sophisticated bearing computer programs. Material and lubricant 

selection must be integrated into the bearing deslgn. Bearing thermal management must be implemented through proper lubricatlon and cooling. 

However, hjgh-speed bearings 

Design parameters must be carefully chosen and optimized 

BIBLIOGRAPHY 

Bamberger, E.N., Averbach, B.E., and Pearson, P.K., Improved Fracture Toughness 

Bearings, AFWAL-TR-84-2103, Jan. 1985. 

Bamberger, E.N., Zaretsky, E.V., and Signer, H., Endurance and Failure Characteristic of Main-Shaft Jet Engine Bearing at 3x106 DN, J. Lubr. Technol., 

Vol. 98, NO. 4, OCt. 1976, pp. 580-585. 

Bamberger, E.N. et al., Life Adjustment Factors for Ball and Roller Bearings - 

An Engineering Design Guide, ASME, N.Y., 1971. 

Fischer, G.K., ed., Advanced Power Transmlssion Technology, NASA CP-2210, 1983, 

pp. 193-269. 

Greby, D.F., What Turbine Technology is Teaching us About High-speed Roller 

Bearings, Mach. Des., Vol. 42, No. 11, Apr. 20, 1970, pp. 229-234. 

Hamrock, B.J., and Anderson, W.J., Rolling-Element Bearings, NASA RP-1105, 

i 983. 

Harris, T.A., Rolling Bearing Analysis, 2nd ed., John Wiley 8 Sons, N.Y., 

1984. 

Parker, R.J., Pinel, S.I., and Signer, H.R., Performance of Computer-Optimized - 

Tapered-Roller Bearings to 2.4 Million ON, J. Lubr. Technol., Vol. 103, No. 1, 

Jan. 1981, pp. 13-20. 

. 

Schuller, F.T., Operating Characteristics of a Three-Piece-Inner-Ring LargeBore Roller Bearing to Speeds of 3 Million ON, NASA TP-2355, 1984. 

Zaretsky, E.V., Schuller, F.T., and Coe, H.H., Lubrication and Performance of 

High-speed Rolllng-Element Bearings, AXE Preprint 85-AM-3F-2, May 1985. 

92 
1ABLE 1. - COMPUTER OMTLMIZED DESIGN FOR 

HIGH-SPEED TAPERED-ROLLER BEARING 

Cup half angle ............ 15'53' 

Roller half angle .......... 1'35' 

Roller large end diameter, m (In) . . 18.29 

(0.720) 

Number of rollers ............ 23 

(1.3456) 

(6.105) 

Bearing outslde diameter, mm (in) . . 190.5 

Total roller length, mm (In) ..... 34.18 

Pitch diameter, m (in) ....... 155.1 

Roller crown radius, m (In) .... 25.4~10 (7.500A 

( 1000) 

percent of apex length ........ 80 

Roller spherical end radius, 

TABLE 2. - PROPERTIES OF TtlHAESILR LUBRICAN1 

Additives ............ Antiwear, oxldatlon 

inhibitor, antifoam 

311 K (100 OF) ................ 28.5 

Kinematic viscoslty, cS, at: 

372 K (210 OF) ................ 5.22 

477 K (400 OF) ................ 1.31 

Flash point, K (OF) ........... 533 (500) 

Pour point, K (OF) ............ 214 (-75) 

(400 OF)), ut X ............... 3.2 

Autoignltion temperature, K (OF) ..... 694 (800) 

Volatllity (6.5 h at 477 K 

Specific heat at 477 K (400 OF), 

J/( kg) (K) (Btu/( 1 b) ( OF) ) ....... 2340 (0.54) 

Thermal conductivity at 477 K 

(Btu/(h)(ft)(OF) ........... 0.13 (0.075) 

Specific gravity at 477 K (400 OF) ...... 0.850 

(400 OF) , J/(m)(s)(K) 
3. !&lo6 

r 

z 

n 

t e- 20 w 

VI 

n 

1.5 

a 

m 1.0 

U 

W 

~64 TF30 CF6 

J75 Od5?30A& 

A A A JT8a 

J57 oJT30 

Jn J85 

HOUSING , PLATE 

?SUPPORT, 

=,‘ 1 BEARI ING /’I 

I I O1 

. 21 41 0 AMBIENT 

AIR (KNOWN 

TEMPERATURE) \ 

RIG UPPER 

HOUSING PLATE 

3 ‘1 I 

P 37 t 

I 

I 

t 

I 

‘%JPPORT 

I BEARING 

TEMPERATURE) 

Figure 2. - Nodal system used for thermal routines in Cybean. 

94 
--- 

(a) Free-contact 

angle (no load). 

I I 

-I-- 

A -- WI 

(b) Contact angle (c) Contact angles 

when under load. when under load 

and at high speed. 

Figure 3. - Changes in contact angle with load and speed. 

(a) Radial load. 

with two-point preload. 

(b) Radial load and two-point preload. 

Figure 4 - Roller bearing load distribution for simple radial load and 

95 
c 

_-------.---.---.-- SIDE 

DIAMETER 

I 

I 

RADIUS 

CORNER RADIUS 

CROWNED ROLLER 

Figure 5. - Crowned roller for high-speed roller bearing application. 

GUIDE 

I I I I I 

0 2 4 6 

CROWN DROP, in 

Figure 6. - Effect of crown drop on skewing moment for 

various amounts of misalignment 

96 
. 

OUTER RACE 

\ c 1 

INNER RACE 

Figure 7. - Three piece inner-race roller bearing. 

FLAT END L.--CONTOURED END 

~NGE 

N 

/ I -\ 

‘y’ 

----7------ 

‘CONTACT 

REGION 

RING 

(a) Conventional, 

one-piece race. 

inner race 

bearing. 

(b) Modified configuration, three-piece inner 

Figure 8. - Roller end-contact geometry of one- and three-piece inner-race 

97 
OUTER RING - . 

0 

0 

/ 

INNER RING-/ 

OUTER RING I, 

\ 

c 

JET 1 

I 

0 --- 

INNER RING 1 

/’ ‘ 

/ 

CONE J’ 

Figure 9. - Placement of jets for ball bearings with relieved 

rings and tapered-roller bearings. 

98 

. 
7 OIL DISCHARGE // '\ 

LI 

\LUNDER RACE 

COOLING GROOVES 

(a) Cylindrical roller bearing. / 

/- '. / 

01 L DISCHARGE v 

ET 

COOLING HOLES 

(b) Ball thrust bearing. 

Figure 10. - Under-race lubrication system for main shaft bearings 

on turbofan engine. 

OUTER RACE 

ROLLER 

LI GHTWEI GHT 

.- . . . . . . . . SEPARATOR 

INNER RACE 

LUBRICANT FLOW 

KUBRICANT 

SUPPLY HOLES 

Figure 11. - High-speed roller bearing with under-race lubrication and cooling. 

99 
01 L 01 L 

OUT IN OUTER RING 

COOLING 

ANNULUS -. 

\ -\ \ / 

'4 \ 

 LUBRICANT \ 

\ 

SUPPLY HOLES SHAFT 

(a) Angular-contact ball bearing. 

CUP 

COOLING 

OUT IN 

tI 

'\ GNNULAR 

GROOVES 

1 rLUBRlCATlON TUBE t I 

0) Computer optimized tapered-roller bearing. 

Figure 12. - Outer-race cooling with under-race lubrication for 

h ig h -s peed bearings. 

100 c --A 
LUBRICATION AND COOLING FOR HIGH SPEED GEARS* 

Dennis P. Townsend 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

The problems and failures occurring with the operation of high speed gears 

are discussed. The gearing losses associated with high speed gearing such as 

tooth mesh friction, bearing friction, churnlng, and windage are discussed with 

various ways shown to help reduce these losses and thereby Improve efficiency. 

Several different methods of oil jet lubrication for high speed gearing 

are given such as lnto mesh, out of mesh, and radial jet lubrication. The 

experiments and analytical results for the various methods of oil jet lubrication are shown with the strengths and weaknesses of each method discussed. 

The analytical and experimental results of gear lubrication and cooling 

at various test conditions are presented. 

need of improved methods of gear cooling at high speed and high load conditions. 

These results show the very definite 

INTRODUCTION 

There are many applications where gears must operate at high speeds 

>50 m/s (10 000 fpm) and sometimes at high tooth loading. Some of these applications jnclude turbine powered ship propulsion, surface effect ships, turboprop, V/STOL aircraft, and geared turbofans. There is very little information 

available on lubrication and cooling methods, or methods to determine the power 

losses and lubrication requirements for gears operatlng at these high power 

conditions. Many high speed gear sets today are operating at load conditions 

below those that could be obtained with better cooling methods. Also many gear 

boxes could operate at better e f f i c i e n c i e s w i t h Improved lubrication and 

cooling methods. 

There are several methods used for lubrlcatlon and cooling gears. Some 

of these methods Include splash lubrlcatlon, drip feed, grease, oll mlst, and 

low or moderate pressurized oil jet flow. Some gear boxes are operated with a 

given lubrication method such as low pressure oil jets with lnto mesh and/or 

out of mesh simply because It has worked in the past and no effort has been 

made to improve the system for better efficiency or hlgher power rating. 

For gears operating at low speeds and low loads the lubrication method is 

not critical since very little heat is generated. Under these conditions the 

gears require lubrication but very little coollng. When the heat input to the 

gear teeth is increased by high speed and high power conditions then the major-' 

ity of the lubricant must be used for cooling the gear teeth. When inadequate 

cooling is supplied to the gear teeth several things may happen. The gear 

teeth can suffer what is called micropitting or grey staining (fig. 1, ref. 1) 

which Is caused by insufficient oil fllm between the teeth. The gears can have 

early fatigue spalls (fig. 2, ref. 2) resultlng from reduced hardness caused by 

'Prepared for the Orlglnal Equipment Manufacturlng Deslgn Conference, Phlladelphla, Pennsylvanla, September 9-11, 1985 (NASA TM-87096). 

101 
a temperature rise of the gear teeth above the gear tempering temperature. The 

gears can fail by scoring (fig. 3, ref. 3) as a result of a loss of the elastohydrodynamic EHD or extreme pressure EP boundary film. The gears can also fail 

by tooth breakage as a result of reduced strength when the teeth are overheated 

and lose their hardness. The objective of the work reported herein is to show 

the gear designer the various lubrication methods that are used In gearing and 

what can be expected with each method, and to provide the gear designer with 

methods to reduce gear losses and thereby reduce heating and improve 

efficiency. 

NOMENCLATURE 

22 A bearing area, cm (in ) 

a addendum, mm (in) 

B 

C 

P 

cs 

D 

d 

di 

dm 

F 

fO 

fl 

hl 

LP 

M 

M1 

backlash, mm (In) 

specific heat, cal/g OC (Btu/lb OF) 

basic static capacity, N (lb) 

root diameter, mm (in) 

gear diameter, mn (in) 

impingement depth, mm (in) 

bearing mean diameter, mn (in) 

face width, mm (in) 

bearing axial load, N (lb) 

bearing load, N (lb) 

bearing radial load, N (lb) 

static equivalent load, N (lb) 

friction factor viscosity 

friction factor load 

bearing clearance, mm (in) 

distance from Impingement point to center line of gears, mm (in) 

tooth module, mm 

bearing friction torque, N-m (lb-in) 

bearing load friction torque, N-m (lb-in) 

102 
Greek symbols 

~ 

U 

M 

m 

N 

P 

I'd 

AI' 

9 

Q 

R 

r 

r 

0 

AT 

V 

V 

V 

9 

j 

W 

bearlng vlscous frlctlon torque, N-m (lb-In) 

gear ratlo 

speed, rpm 

power loss, kW 

dlametrlal pltch, l/ln 

011 jet pressure, N/cm (lb-in) 

heat loss, cal/mln (Btu/mln) 

pitch radius, m (In) 

radius to 011 Jet Implngement, m (In) 

gear outslde radius, mm (In) 

oil jet offset, m (In) 

oil temperature rise, "C (OF) 

veloclty, m/sec (ln/sec) 

gear pitch llne velocity, m/sec (ft/sec) 

011 jet velocity, m/sec (ft/sec) 

oil flow, g/mln (lb/ln) 

a 

R 

Y 

eW 

x 

11 

U 

0 

4 

+!3 

9 

W 

bearlng contact angle, degrees 

011 jet angle, radians 

nondlmenslonal Impingement depth = dl/whole depth 

gear rotatton angle for 011 jet in tooth space, radlans 

gearbox space functlon l-free space, 0.6-0.7 for large enclosure, 0.5-0.6 

for fitted gear case 

vi scosl ty, reyns 

lubricant vlscoslty, centlstokes 

oil mixture functlon l.O=oll free atmosphere 

gear pressure angle, radians 

gear rotatlonal speed, radlans/sec 

103 
POWER LOSSES 

There are f0u.r main areas of losses in high speed gears. These losses 

can be controlled for the most part by careful design and construction. The 

losses consist of bearlng losses, tooth friction losses, oil churning losses, 

and gear windage losses. 

losses particularly when fluid fllm bearlngs are used. However, when long llfe 

is a requirement, the fluid film bearing is the best choice. An approximate 

method for the friction torque in a fluid film bearlng Is shown in the following equation from reference 4. 

Bearing losses may account for about half of the 

Rolling contact bearings have conslderably less friction loss when 

properly lubricated and wlll generally have a much shorter life than a fluld 

film bearlng. However, rolling element bearings should have sufftclent life 

for many appllcatlons. The friction torque for rolling element bearlngs may 

be estimated by the following equations from reference 5. 

Mb = M1 + M 

U 

The friction torque due to the applled load is 

= F if larger r 

Fa I 0.9 Fa cot a - 0.1 Fr or Fa 

H,=fFd 1 am 

Y 

, z = 0.001, y = 0.33 FS 

where f1 = 

for 30 angle contact bearlngs and f1 = 0.0003 for roller bearlngs. 

cous frlctton torque may be estimated by 

The vis- 

-5 2/3 M = 1.42~10 fo(VoN) dm3 

U 

M = 3.492~10-~ fo dm3 uoN 5 2000 U 

for ball bearlngs fo values range from 3 to 4. Bearing computer programs 

have been developed that give more accurate results for bearlng torque, 

(refs. 6 and 7). 

The tooth friction loss is probably the lowest loss in the gear system 

when the gears are adequately lubricated. There Is very little that can be 

done to reduce gear tooth friction loss once adequate lubrlcatlon has been 

provided. Some lubrlcants will have a little less friction loss than others. 

Some gears will have more or less tooth frictlon loss because of the type of 

design mainly because of the different slldlng conditions. For Instance, high 

contact ratio spur gears generally have more slidlng and, hence, more losses 

than standard contact ratlo gears. .However, slnce thls 1s a small part of the 

overall loss, it generally has little effect on the total loss. 

104 
Windage losses can account for a large part of the total gear box losses 

ln high speed gears because of the high pitch line velocities. Some of this 

wlndage lose can be reduced by careful designs. for Instance, it was shown In 

reference 8 that axlal holes in a gear web can significantly tncrease the wtndage losses. Also it was shown that placing a shield on the ends of the gear 

teeth, to prevent the air circulating into the teeth, reduced the windage 

losses by a large percentage. The wlndage losses in a gear box wlth the smooth 

sides of the box located approximately 1 In from the gear and the Inside diameter 0.6 In from the teeth, reduced the wlndage loss to approxlmately one-half 

that for open gears. The pumping loss of the air at th-e entrance of the mesh 

accounts for some losses and can be reduced by reducing the pressure in the 

gear box which also reduces other wtndage losses. An equatlon for approxlmating the windage loss in gears was glven In reference 8 as shown in the follow1 ng equation. 

Sides Perfphery 

(4) 

2.9 F0.75 1.15 P = N~*~(o.I~ 03a9 + D M ) x lo-*'. 0.x 

Churning losses are caused by the gear striking, pumping, or otherwlse 

moving the lubricant around In the gear box. It is very important In high 

speed gear boxes to get the lubricant to perform its lubrication and cooling 

function and then get it out of the way. Shrouds are sometlmes used ln gear 

boxes to direct the lubricant oil away from the gears.. If too much lubricant 

is allowed to enter the gear mesh, excesslve losses will'occur from oil belng 

trapped in the gear teeth and belng pumped out of the mesh in the axlal direction. In spur gear application this is more critical than helical gears. This 

is one reason why most gear designers prefer helical gears for high speed gearing. Even with helical gearing, however considerable power loss occurs wlth 

too much oil getting Into the gear mesh. Some spur gear designs have a groove 

In the center of the axial length of the gear to reduce pumping losses ln the 

mesh. This Is also done in some very wide helical gear applications. 

LUBRICATION AND COOLING METHODS 

There are many high to moderately loaded high speed gears operating today 

wlth oil jet lubrlcation uslng 30 to 50 psig oil pressure to lubricate the 

gears. This type of low pressure system does not do a good job of cooling the 

gears in a high speed gear drive and will only allow the gears to operate at a 

moderate load. In a low pressure oil jet system the oil jet can only penetrate 

a small distance into the tooth space. Thls results In cooljng of the tips of 

the gear teeth only. This causes the gear tooth temperature to be higher than 

that obtained with a better system, such as high pressure radial oil jet. 

figure 4(a) is a calculated temperature profile of a gear tooth cooled with low 

pressure operating at a moderate speed and high load reference 9. When the 

speed is increased at this load condition and low pressure lubrlcation, fallure 

of the gears will generally occur. Flgure 4(b) shows how the gear tooth temperatures are reduced when the oil jet pressure Is increased to obtain good 

implngement depth. 

OUT OF MESH JET LUBRICATION 

A large number of gears are lubrtcated wtth low pressure oll jets into 

mesh or out of mesh or both. In the out of mesh lubrication method the 011 

105 
jet has a very modest Impingement depth. Thls Is Illustrated In flgure 5 

whlch shows the analytlcal results for Impingement depth on the pinion uslng 

the following equation from reference 10. 

dl = ro - r2 + LS (5) 

Figure 6 Is a hlgh speed photograph of an oil jet at the out of mesh condltlon. In order to get the maxlmum Impingement depth for the out of mesh 

condition care must be exerclsed to get the proper 011 jet location. The analysls Indicates that the plnlon can be completely mlssed by a very small change 

In offset distance from the lntersectlon of the outslde diameter of the gear 

and pinion or from a small change In the jet angle. For maxlmum Impingement 

depth In most cases the 011 jet should be directed at the lntersectlon of the 

two outslde diameters at an angle that will intersect the pitch point of the 

gear and plnlon (ref. 10). For large gear ratios It Is probably better to 

favor the pinion to get a better coollng balance. Reference 11 glves an analytlcal method for out of mesh jet lubrication for gears wlth modifled center 

distance and/or addendums. Reference 11 also gives the lmplngement depth 

results for varlous 011 jet off-sets and 011 jet angles. Figure 7 shows the 

analytlcal results for the out of mesh jet lubricatlon for varlous 011 jet 

angles . 

INTO MESH JET LUBRICATION 

Into mesh 011 jet lubrlcatlon Is often used as a means of getting oil to 

the gear tooth surfaces at a good Implngement depth when the 011 system Is 

operatlng at a low pressure. Thls method is effective because It uses the 

gear tooth velocity movlng wlth the 011 jet velocity as shown In flgure 8. 

References 12 and 13 glve equatlons for the 011 jet Impingement depth for Into 

mesh lubrication. When the jet veloclty Is less than the gear velocity, the 

011 Impinges on the backside of the teeth as shown In figure 9. When the jet 

veloclty Is greater than the gear velocity, the oil will lmplnge on the front 

of the gear tooth. Thls Is shown In flgure 8. The optimum lmplngement depth 

for into mesh lubrlcatlon was shown In references 12 and 13 to occur when the 

011 jet veloclty and gear veloclty are equal. The equation for 011 jet 

Impingement depth at the oplmum velocity Is given by the following equatlon. 

dl = 1/Pd = a Vj = wg Rsec 0 (6) 

When Into mesh lubrication Is used wlth high speed gears care must be taken to 

avold excesslve 011 belng trapped In the gear teeth. This trapplng can cause 

varlous problems such as loss of efflclencles, hlgh loads on the teeth, hlgh 

noise, and even gear failure under some condltlons. In many cases the bulk of 

the coollng 011 Is supplied to the out of mesh location wlth only a small 

percentage for lubrication supplied to the into mesh posltlon reference 14. 

However, there Is usually sufficient 011 film remalnlng on the gear tooth for 

good lubrlcatlon when adequate cooling is provided at the out of mesh locatlon. 

In some cases where It Is dlfflcult to keep the oil out of the Into-mesh zone, 

a clrcumferentlal groove will be cut Into the center of one of the gears to 

break up the length of the teeth; thereby, reducing the trapplng losses. This 

groove reduces the axial length requlred to pump the 011 by at least one-half. 

106 
RADIAL JET LUBRICATION 

When the oil jet Is directed radially Inward the best impingement depth 

Is obtained. 

on the face of the tooth, the radial oil jet offers the best method of gear 

lubrication and cooling. Figure 10 Is a high speed photograph of the radial 

directed 011 Jet and shows the 011 jet penetrating the tooth space just before 

tmplngement on the gear tooth flank.. The 011 pressure here Is sufficient to 

allow the oil jet to Impinge on the gear tooth a little more than half way 

down the working depth of the gear tooth. The maximum cooling Is obtained 

when the oil pressure is suffictent to cause the oil jet to reach an Implngement depth equal to the full worklng depth of the tooth. However adequate 

coollng can often be obtained with Impingement depth just below the pitch llne. 

When radial Jet lubrlcatlon is used the 011 jet should be located near the out 

of mesh position wlth the jet directed radially at the center of the gear and 

plnion. In a speed reducer the pinion will receive cooling on the loaded side 

of the tooth while the gear will be cooled on the backface of the tooth. When 

the gear set Is a speed increaser, the pinion will receive cooling on the backface of the tooth and the gear on the loaded side. Experiments have shown 

(ref. 9) that good coollng of the gear or pinion can be obtained when either 

the loaded flank or unloaded flank of the gear tooth Is cooled. Flgure 11 

shows the effect of oil Jet pressure and load on gear tooth temperature using 

radlal 011 jet cooling on the back flank of the gear tooth. The temperature 

was measured on the loaded flank of the gear tooth. Thls figure very clearly 

shows that good cooling Is obtained when coollng the back flank of the tooth. 

depth on the tooth flank for various speeds and oil jet pressures for a radial 

Since gear tooth cooling is a maxlmum when the oil jet implnges 

The following equatlons from references 9 and 15 gives the Impingement 

directed oil jet. i 

1.5708 t 2 tan 4 t 8/2 

dl =N.I (7) 

t tan + u 

pd 2977 Ap 

The vectorial model used to calculate the radial lmplngement depth Is shown In 

figure 12. The analysis and experimental results for a radial directed 011 

jet are shown In figure 13. When the impingement depth for a glven gear operatlng condltlon Is known or desjred then the pressure requlred to obtain that 

Impingement depth Is glven by the following equation. 

2 

dl 'd Nd 

I" 

(2 - d,P,) tan 4 

When the oil jet velocity equals the gear veloclty the oil jet will usually Implnge to a depth approximately equal to the full worklng depth and 

provide very good coollng for the gear teeth. For standard gear tooth geometry 

the pressure requlred to obtaln thls velocity may be approxlmated by the followlng equation where V is the m/sec (ft/sec) and ~p Is In N/cm2 (psi) 

107 
2 

2 

Ap = Vg /169 English 

Ap = Vg /22.8 Metric (9) 

Using the above equation for gear operating at a pitch line velocity of 

150 m/s (500 ft/s) the oil jet pressure required for full depth impingement 

would be approximately 1014 N/cmZ (1480 psi). This pressure is much higher 

than the o i l j e t pressure used in most high speed gear boxes operating at 150 

to 300 m/s (500 to 1000 ft/s). These gear boxes must operate at reduced loads 

because for the limited cooling available. It should be understood that the 

oil jet site must be reduced at these high pressures to limit the oil flow to 

that required for good cooling. Using the minimum orifice size of 1.02 mm 

(0.04 in) specified by many gear designs would give an oil flow of approximately 1.3 gpm per orifice which may be too much oil for most applications. 

However if the orifice size Is reduced to 0.5 mm (0.02 In) then the oil flow 

Is only 0.33 gpm or 1/4 of that for the larger orifice. 

orifice plugglng the oil should be filtered through a 5 or 10 pm filter. It is 

much better for both the gears and bearings to filter the oil through a 5 pm 

or better filter. The finer filteration will also improve the gear and bearing 

life. 

In order to limit 

COOLING REQUIREMENTS 

This amount of lubricating fluid required for coolIng of gears and bearings may be determined by estimating the power loss for gears, beatings etc., 

and then using an appropriate temperature rise in the oil, the required oil 

flow can be determined. The following equatlon can be used to determine the 

oil flow in a given gear system. Assume 2000 hp is transmitted in a gear set 

with one mesh and two sets of bearings. The losses per mesh In a well designed 

spur or helical gearbox should be no more than 0.5 percent with the losses 

equally divided between the gears and bearings. 

0.005 x 2000 or 10 hp, which is 424.4 Btu/mln. Assuming a 50 OF temperature 

rise, the total oil flow required would be, 

w = Q/cp~T = 424.4/0.6 x 50 = 14.2 lb/mln 

(10)with equal flow to the gears and bearings of 7.1 lb/min. If the gear set 

ls less efficient than the above then more oil flow would be required. 

The oil therefore must absorb 

SUMMARY AND CONCLUSIONS 

There are several types of losses In gearbox systems that should be evaluated in a high speed gear design. These losses include the gear tooth sliding 

and friction, the windage of the air flowing around the gears, the churnlng 

losses of the oil being pumped or accelerated by the gears and the bearing 

losses of rolling element or fluid film bearings. Good design practices can 

reduce the effect of these losses. 

The gears may be lubricated by one or more of several methods. For high 

speed gears the pressure o i l j e t is definltely required to provlde good cooling. Oil jets may be directed into mesh, out of mesh, or In a radial direction 

at various oil jet pressures. However, for the maximum cooling and lubrlcation 

of the gear set the o i l j e t should be directed radially at the gear and pinion 

near the out of mesh posltlon. This will provide good coollng and lubrication 

108 
an keep the 011 from entering into mesh zone where excessive losses may occur. 

The conclusions may be summarized as follows. 

(1) Gearbox efflclency for hlgh speed gearlng can be improved by careful 

attentjon to reducing windage and churning losses. 

(2) The best method for lubrication and cooling of hlgh speed gearing is 

a hlgh pressure 011 jet at the out of mesh posltion directed radjally at the 

gear and pinion with sufficient velocity or pressure to Impinge on the tooth 

flank well below the-pitch line. 

at conslderable higher power levels than the majorlty of hlgh speed gearboxes 

are operatlng at today. 

c 

(3) When the above condltions are performed, high speed gears can operate 

REFERENCES 

Internatlonal Symposium on Gearing and Power Transmissions, 

Sept. 3, 1981, Tokyo, Japan, vol. 1, pp. 319-324. 

2. Townsend, Dennis P.: The Application of Elastohydrodynamic 

in Gear Tooth Contacts. NASA TH-X-68142, 1972. 

3. Townsend, Dennis P.; Baber, Berl 8.; and Nagy, Andrew: Eva 

1. Dudley, Darle W.: Characteristics of Regimes of Gear Lubrlcatlon. 

AUg. 30- 

Lubrl cat I on 

uation of H 

Contact Ratio Spur Gears With Profile Modification, NASA TP-1458, 1979. 

4. Fuller, Dudley, D.: Theory and Practice of Lubrlcatlon for Engineers. 

Wlley, 1965. 

5. Harris, Tedric A.: Rolling Bearlng Analysls. Wlley, 1966. 

6. Pirvics, Juris; and Klechner, Robert J.: Predlcatlon of Ball and Roller 

. Bearing Thermal and Kinematic Performance by Computer Analysls. Advanced 

Power Transmlsslon Technology, NASA CP-2210, G.K. Fjschet, ed., 1983, 

pp. 185-201. 

7. Coe, Harold H.: Predicted and Experimental Performance of Large-Bore 

High-speed Ball and Roller Bearings. Advanced Power Transmlsslon 

Technology, NASA CP-2210, G.K. Flscher, ed., 1983, pp. 203-220. 

8. Dawson, F.H.: Wlndage Loss In Large Hlgh-Speed Gears. Proceedings of the 

Instltutlon of Mechanical Engineers, Part A - Power and Process Englneerlng, vol. 198A, no. 1, 1984, pp. 51-59. 

9. Townsend. D.; and Akln. L.: Analytlcal and Experlmental Spur Gear Tooth 

Temperature as Affected by Operatjng Variables. J. Mech. Des., vol. 103, 

no. 1, Jan. 1981, pp. 219-226. 

10. Townsend, D.P.; and Akln, L.S.: Study of Lubrlcant Jet Flow Phenomena in 

Spur Gears - Out of Mesh Condition. J. Mech. Des., vol. 100, no. l., 

Jan. 1978, pp. 61-68. 

109 
11. 

12. 

13. 

14. 

15. 

Akin, L.S.; and Townsend, D.P.: Lubrication Jet Flow Phenomena In Spur 

and Helical Gears with Modified Center Distances and/or Addendums. For 

Out-of-Mesh Conditions. Journal of Mechanisms, Transmissions, and Automation in Design, vol. 107 no. 1, Mar. 1985, pp. 24-30. 

Akin, L.S.; and Townsend, D.P.: Into Mesh Lubrication of Spur Gears with 

Arbitrary Offset 011 Jet. Part I - For Jet Velocity Less Than or Equal 

To Gear Velocity. Journal of Mechanisms, Transmissions and Automation in 

Design,’ vol. 105, no. 4., Dec. 1983, pp. 713-178. 

Akfn. L.S.; and Townsend, D.P.: “Into Mesh Lubrication of Spur Gears 

with Arbitrary Offset Oil Jet. Part I1 - For Jet Velocities Equal to or 

Greater Than Gear Velocity. Jounral of Mechanlsms, Transmissions and 

Automation in Design, vol. 105, no. 4., Dec. 1983, pp. 719-724. 

Tucker, A.I.: Bevel Gears at 203 m/s (40,000 FPM). ASME Paper 

77-DET-178, Sept. 1977. 

Akin. L.S.; Townsend, D.P., and Mross, J.J.: Study of Lubricant Jet Flow 

Phenomena in Spur Gears.. J. Lubr. Technol., vol. 97, no. 2, Apt. 1975, 

pp. 283-288. 

110 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

I- 

(a) Spur, 60HRC. 

(bl Helical, 60HRC. 

Figure 1. - Examples of micro pitting. 

111 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

Figure 2. - Destructive pitting: Heavy pitting has taken place, predominantly 

in thededendum region (ref. 2). 

(al Standard gear. (b) New-tooth-form gear. 

Figure 3. - Typical scoring failures. 

112 
tal Zero impingement depth. 

11 K 

(bl 87.5 percent Impingement depth. 

Figure 4 -Calculated gear tooth temperatures speed 10 3.N rpm. load 5903 Nlcm (3373 Iblin). 

113 
ORlGlNAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

1.5 1 112 113 114 115 116 

Figure 5. - Gear ratio vs nondimensional impingement depth speed 36603 rpm jet 

pressure 17~10~ Nlm2 (25 psi) 28 pinion teeth. 

(a) Oil jet clearing pinion teeth. 

8. 3x104N/m21 12 psi). 

ID1 UII jet clearing qear teeth. 

Figure 6. - Oil jet impingement depth, out of mesh. Speed. 3600 rpm; jet pressure, 

114 
.MIORlGINAL PAGE: 

BLACK AND WHITE PHOTOGRAPH 

3 250 500 loo0 1500 2oM) 

JET VELOCITY. Vi. inlsec 

Figure 7. -Effect of pi on impingement depth, Si = 1.0, 

21/35 gear ratio. 

Pi 

-1 

-. 5 

0 

OPT 

.5 

1.0 

Figure 8. - Into mesh lubrication for Vj >Vg 

115 
ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 

Figure 9. - Into mesh lubrication for V, >Vy 

Figure 10. -Radial jet lubrication so00 RPhloil jet pressure 21N/cm2 MPSI. 

116 
OIL JET PRESSURE 

0 1&104Nlm2(20psi) 

A 41~14~Nlm~(60psi) 

0 97x104Nlm2(140 psi) 

4m- 

'& 360 W I 4 160 

120 

80 

320 - 

1 I 

2030 3000 4000 5000 6000 

Nlcm 

I I I 

loo0 1500 Moo 2500 3a30 3500 

lblln 

Figure 11. - 1. R. microscope measurement of gear average surface temperature vs 

loll for three oil jet pressures, speed 7Hn rpm, oil jet diameter a04cm (a 016 in) 

inlet oil temperature 308 K 195 9). 

--I----- 

TAN 0 

'd 

-l/P,j(II/2+2TAN@ +a1214 

Figure 12. -Vectorial model for penetration depth. 

117 
CALCULATED 

EXPERIMNTAL 

- 

---- 

0 10 20 40 

PRESSURE, WmZ 

Figure 13. -Calculated and experimental impingement depth vs oil jet pressure at 4920 and 

2560 rpm. 

118 
A Historical Perspective of Traction 

Drives and Related Technology* 

Stuart H. Loewenthal 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

Traction or friction drives are perhaps the simplest of all rotary mechanisms and yet relatively 

little is known and even less has been written about them. In its simplest form a traction drive is just 

two smooth, unequal sized wheels in driving contact. Their simplicity suggests that their existence 

predates that of the gear drive. As speed regulators, oil-lubricated traction drives have been in 

industrial service for more than SO years; yet the concept of transmitting power via traction remains 

unfamiliar and even alien to many. Indeed, traction drives are commonplace in our daily existence. 

Our car tires engaged against the road surface or a locomotive’s driving wheels against the rail are but 

two common examples. 

Many applications may be found in equipment where simple and economical solutions to speed 

regulations are required, such as phonograph drives, self-propelled lawnmowers, and even the 

amusement park ride driven by a rubber tire. Of course, in these examples, simple dry contact is 

involved. However, this same principle can be harnessed in the construction of an oil-lubricated 

transmission containing all steel components. 

Traction drives can be constructed to give a single, fmed-speed ratio, like a gearbox or, unlike a 

gearbox, a speed ratio that can bc continuously varied. This latter arrangement is of extreme interest 

to drive-train configurators since it provides them with an essentially “infinite” number of shift 

points to optimize the performance of their drive system. 

Because power transfer occurs between smooth rolling-bodies, generally across a thin, tenacious 

lubricant film, traction drives possess certain performance characteristics not found in other power 

transmissions. Traction drives can be designed to smoothly and continuously vary the speed ratio 

with efficiencies approaching those of the best gear drives. Unlike transmissions with gear teeth, 

which, even when perfectly machined, generate torsional oscillations as the load transfers between 

teeth, power transfer through traction is inherently smooth and quiet without any “backlash.” Film 

trapped between the rollers, tends to protect against wear and to dampen torsional vibrations. The 

operating speed of some traction drives is limited only by the burst strength of the roller material and 

the available traction in the contact. In many cases traction drives can be designed to be as small as or 

smaller than their nontraction-drive counterparts. When manufactured in sufficient quantity, costs 

can also be quite competitive because of the similarities in manufacturing tractiondrive components 

and ordinary mass-produced ball and roller bearings. 

Although traction drives have been available for some time (refs. 1 to 7). it is perhaps within the 

past IS years or so that they have been considered serious competitors of conventional mechanical 

power transmissions in some of the more demanding applications. The earlier drives, particularly 

those targeted for automotive applications, had their share of durability problems above nominal 

power levels and, as a consequence, relatively few succeeded in the market place. The underlying 

reason for this was that certain critical pieces of technology were generally lacking. Designs were 

based on mostly trial and error. No uniform failure theories were available to establish service life or 

reliability ratings. The drive materials of the day were crude by today’s standards. In short, traction 

drives were in their technical infancy. 

Prompted by the search for more efficient automotive transmissions and bolstered by advancements made in rolling4ement bearing technology, interest in traction drives has been renewed. 

Today’s analytical tools, materials, and traction fluids are far superior to those available only 

10 years ago. This has lead to the reemergence of traction drives and the technology related to their 

design. 

It is the intent of this review to trace the evolution of tractiondrive technology, in a limited 

sense, from its early development to the efforts underway today. The reader will appreciate that no 

attempt has been made to comprehensively document the history of traction drives but rather to 

*Published in Advanced Power Transmission Technology.” NASA CP-22 IO. 1981. 

119 
I 

Figure 1. - Assyrians positioning stone figure using log rollers and “human 

traction,’ circa 700 B.C. (ref. 8). 

assemble, in a single place, a collection of source material on this broad subject for the benefit of 

those unfamiliar with these interesting devices. 

Early History 

Since the development of machines there has been a need for transmissions to effectively match 

the speed of the power source to that of the driven device. In early human history, the power sources 

were either human or animal and in “direct” drive with the load. This is vividly illustrated in figure 1 

(ref. 8). This scene (ca. 700 B.C.) depicts Assyrians positioning a stone figure with the aid of log 

rollers and “human traction.” Traction as an assistance to motion is a natural part of the human 

experience. Also depicted are humans pulling wheeledcarts. It is clear that the ancients had 

knowledge of both the roller and the wheel. Thus, the ingredients to develop simple traction wheel 

mechanisms were available to them. 

The existence of mechanical drive systems may even predate this by hundreds of years, according 

to Dudley (ref. 9). He speculates that gear devices might have been used in clocks, temple devices, and 

water-lifting machines as far back as IO00 B.C. by the ancient Greek, Egyptian, and Babylonian 

civilizations. Unfortunately, there is no tangible evidence to support this view. Drachmann (ref. 10) 

maintains that the origins of the toothed gear wheel are often erroneously associated with a passage in a 

book by Aristotle. The passage was, in fact, not due to Aristotle but to his school (ca. 280 B.C.), and 

there is no explicit mention of teeth on the wheels. These wheels may just as well have been smooth 

disks in Jiiicionul contact according to Drachmann. Although there seems to be no evidence in the 

literature or artifacts to establish when friction wheels were fvst used, it is reasonable to assume that 

before man went to the trouble of cutting notches in smooth wheels, he probably used them as they 

were to transmit rotary motion. Pins or cogs might have been added to the wheel to overcome 

dimensional variations due to the crude, outef-round. wheels of the day or when the driving load 

simply became too large to be sustained by frictional contact. It is ironic that these very cogs, though 

now more sophisticated than in earlier times, are often the source of many power transmission 

problems. Although the writer’s limited efforts failed to uncover tangible examples of friction drives 

before the modem era, the likelihood of their use in primitive rotary machinery is too great to dismiss. 

One of the earliest known examples of a friction drivel is that patented by C. W. Hunt in 1877 

(fig. 2; ref. 11). This drive had a single, spoked transfer wheel, which was probably covered with 

1The term “friction drive” is normally used to refer to a drive that is nonlubricated. while “traction drive” refers to one with 

oil-wated components. The friction terminology may have evolved from the fact that these drives intentionally used at least 

1 20 
C. W. HUN?. 

Coontrr-flhrft for Drirlog Yaohinory. 

No. 197.472. Patontod Nov. 27,1877. 

FJ:2. Q. 4 

Figure 2. - Hunt’s 1877 toroidal friction drive (ref. 11). 

Figure 3. - 1899 Hoffman toroidal friction drive for belt-driven machinery 

(ref. 12). 

leather, running against a pair of toroidally shaped metal disks. Judging by the pulley flanges 

attached to the toroidal disks, the drive was intended to regulate the speed of beltdriven machinery 

such as that commonly found in factories at the turn of the century. By tilting transfer wheel E, the 

effective rolling radii of toroids B and D could be altered and, hence, their relative speeds. 

A similar drive was devised by W. D. Hoffman as shown in an 1899 British patent application 

(fig. 3 from ref. 12). The toroidal drive arrangement apparently has found great favor with tractiondrive designers through the years. Work continues on this configuration even today, more than 100 

years later! 

one roller that was covered with a high friction material such as leather, rubber, fiber or even wood. This not withstanding, the 

term ‘‘friction’’ is somewhat of a misnomer since it is still the traction force responsible for positive motion of the driven 

element. 

121 
hY 3. twb THE AUTOMOBILE. I906 

rI/ 

We Fd the 

Tra neaiaai on 

k Sbom Harid m lrhdc 

Flywheel end 

Dilea Cllltck 

TheFmmeuof 

Sed. Bakgs. 

kl*dRdk?h 

Material and 

Wortmuhip of 

Hiw Grade 

mea 

F+ sudua 

only in uu om 

Intermediate a u II 

Side Friction 

WH. Do I(rl 

R*l*oaDka 

Drtrr Cmd 

Fnah Wheel in 

Drina om lam. 

medinre Speeds 

by Boll Side Whecl. FTD~ 

Flyhd. Note 

IhDkaQllCL 

in Hub d the 

Flywheel. 

ma4 

ha Lu SPdn 

on the Running 

Gar of 8 Car 

Figure 4. - An early automotive friction drive, circa 1906, (ref. 17). 

Friction drives also found use on several types of wood-working machinery dating back before 

the 1870's. The 1876 edition of Knight's American Mechanical Dictionary (ref. 1) describes a dealframe machine for slitting pine-timber which employed a friction disk drive for regulating the feed 

motion. In another source (ref. 2), a wood panel-planer is described whose feed rolls are driven by 

friction wheels. Appleton's Cyclopedia of Applied Mechanics, published in 1880 (ref. 3). reports of 

frictional gearing being used on wood-working machinery in which one wheel was made of iron and 

the other, typically the driver, of wood or iron covered with wood. For driving light machinery, 

wooden wheels of basswood, cottenwood, or even white pine reportedly gave good results. For heavy 

work, where from 30 to 45 kW was transmitted by simple contact, soft maple was preferred. 

Appleton's Cyclopedia also discusses bevel frictional gearing in which the bevel-gear is a smoothfaced iron cone being driven by a bevel pinion, with a wooden rim. This rim comprises several layers 

of hardwood followed by soft maple bolted on to a flanged-hub made of iron. 

Automotive Transmissions 

It was not until the introduction of the horseless carriage at the end of the 19th century that the 

goal of developing a continuously variable transmission (CVT) for the automobile sparked 

considerable friction drive activity. Although the potential performance benefits associated with a 

CVT for other types of machinery had been recognized before, the safety, simplicity, and reliability 

122 ORIGINAL PAGE - 

BLACK AND WHITE PHOTOGRAPH 
BLACK AND WHITE PHC>TGGiiA?l~f 

of the gearset offered greater appeal. It was good enough to select roughly the right speed ratio and to 

up-size the powerplant slightly, if needed, to drive the load. It quickly became apparent to the 

designers of early automobiles that a highly flexible transmission was badly needed to compensate for 

the lack of flexibility of the early gasoline engines. These engines had a tendency to run well only at 

one speed. As noted by P. M. Heldt in an unusually comprehensive review of the development of the 

automatic transmission (ref. 13),- the chief objection to early sliding-gear transmission, apart from 

their lack of flexibility, was the difficulty in gear shifting. The gearboxes used on many of the early 

vintage cars, such as 1890 Panhard, were adopted from the clash gears used in factory equipment 

(ref. 14). Gear changing was not merely difficult but potentially damaging to the gear teeth. 

According to Hodges and Wise (ref. 15), “the best technique with those early cars was to slip the 

clutch gently and bang the gears home as quickly as possible, in the hope that you might catch the 

cogs unawares.” (Although Prentice and Shiels patented the synchromesh principle back in 1904, it 

wasn’t until the late 1920’s that General Motors developed the syncromesh gearshift for production. 

Syncromesh allowed almost any driver to shift from one speed to the next without clashing the gears 

(ref. 16).) In view of the limitations and inconveniences associated with gear changing, it is not 

surprising that the inventors of the day looked for a simple means of continuously and, hopefully, 

automatically varying the speed ratio between the engine and the wheels. 

Mechanical ratchet, hydraulic, and electro-mechanical drives were all tried, but friction drives, 

-because of their simplicity, were the first automobile transmissions to provide infinite ratio selection. 

The earliest of these was the rubber V-belt drives that appeared on the 1886 Benz and Daimler cars, 

the first massproduced gasolinecngine-powered vehicles. FriAion disk drives, similar in construction 

to the gearless transmission illustrated in a 1906 advertisement (fig. 4 from ref. 17), were used as 

regular equipment on a number of early motor cars. These included the 1906 Cartercar, 1907 

Lambert, 1909 Sears Motor Buggy, and 1914 Metz Speedster. An early advertisement for the 

Lambert appears in figure 5. 

The Cartercar (fig. 6) had an extremely simple friction drive consisting of a metal disk, driven by 

the engine crankshaft, in traction contact with a large, fibercovered spoked wheel mounted on a 

transverse countershaft. The countershaft, in turn, was connected to the rear axle by an ordinary 

chain drive. To vary speed ratio, a driver operated lever (fig. 6) was used to radially position the 

MOTOR AGE MARCH 7, 1907 

DEALERS 

EXAMINE 

THE FR 

USERS 

INVESTIGATE 

CTION DRIVE CAR 

Im 

4 

ka --=a!‘ mm 

L 

THE BUCKEYE MANUFACTURING co., - ANDERSON, IND., u. s. A. bu o..c wu ha ?- L...Y*. !dun- umer~cu Maor car .w=ur.cNnn* A~~OUU~IOQ NOW yon. WM. tr srow A* cu*m 

Figure 5. - An early advertisement for a friction-drive car (ref. 17). 4 

123 
Figure 6. - 1909 Cartercar equipped with a friction, continuously varible 

transmissibn (CVT). (Courtesy of the Henry Ford Museum, Dearborn, Michigan.) 

output follower -wheel across the face of the metal disk-turntable fashion. Neutral was achieved 

when the follower wheel was centered. By moving the follower wheel past center, reverse rotation 

would occur to allow the vehicle to back up. The smoothness and ease of operation of the Cartercar 

transmission made it quite popular. It is not well-known that Mr. W. C. Durant, founder and first 

president of General Motors Company, acquired the Cartercar Company in 1908 because of his 

expectation that friction drives would soon be universally used in automobiles (refs. 16 and 18). In 

1910 the Cartercar Company even produced a Model “T” truck, equipped with their friction drive. 

Despite its catchy slogan, “NO clutch to slip-no gears to strip . . . a thousand silent speeds and only 

one control lever, that’s a Cartercar,” the Cartercar Company’s commercial success was shortlived. 

From 1909 until 1912, Sears marketed a twocylinder, 14-horsepower “Motor Buggy,” also 

equipped with a friction drive (ref. 19). “Absolute simplicity, its positiveness under the most severe 

conditions and its unequalled flexibility,” boasted one of the Sear’s ads. However, by about 1915, 

cars equipped with friction drives had virtually disappeared (ref. 19), presumably due, in part, to the 

need for frequent renewal of the friction material. 

Despite the limited success of these earlier attempts, the goal of designing an automotive 

transmission that smoothly and automatically shifted was not lost. In the late 1920’s the Buick 

Division of General Motors was given the task of developing a continuously variable, oil-lubricated, 

steel-on-steel traction drive. This transmission was similar in design to the Hayes double toroidal 

traction drive, patented in 1929. The Hayes Self-Selector Transmission (ref. 20), although originally 

developed here, was later offered as an option on the 1935 British Austin Sixteen (ref. 21). 

The General Motors toroidal drive, later called the toric transmission, is illustrated in figure 7. 

The geometry of the drive is remarkably similar to the 1877 Hunt drive, with the addition of a second 

toroidal cavity and a ball differential to balance loading between the two cavities. An extensive test 

program was conducted on this drive. Seventeen road-test vehicles equipped with the toric drive 

accumulated over 300 OOO miles of road testing (ref. 18). A 20-percent improvement in highway fuel 

mileage was reported. In 1932 General Motors decided to produce this type of transmission (ref. 16); 

however, no cars equipped with the toric drive were ever sold to the public. The reasons for halting 

production were never really made clear. Some say that there were unresolved discrepancies in 

service-life data obtained during road tests and that obtained from laboratory bench testers. Others 

believe that the availability of premium quality bearing steel, needed in large amounts to make each 

drive, was simply not great enough at the time to meet expected production requirements. Whatever 

the reasons, Alfred P. Sloan, Jr., then president of General Motors, turned the transmission down 

for production in the belief that it would simply be too expensive to make (ref. 16). 

ORIGINAL PAGF 4 

124 

BLACK AND WHITE PHOTOGRAPH 
Figure 7. - General Motor’s Toric transmission (ca. 1928). 

. Work on various types of automatic transmissions, beginning in the early 1920’s finally lead to 

the production of the 1937 Oldsmobile semiautomatic, self-shifting transmission. This transmission 

still required the use of a clutch pedal for stopping and starting. The hydramatic soon followed. It 

was the first mass-produced, fully automatic transmission-first appearing on the 1940 Oldsmobile 

(ref. 16). However, the old toric drive was not completely forgotten, it reappeared briefly in the 

1960’s on General Motor’s experimental gas turbine RTX bus. 

General Motor’s New Departure Bearing Division produced an industrial counterpart to the 

toric drive. By 1935 when production was halted, over 1600 units of the “Transitorque” traction 

CVT had been marketed (ref. 20). The drive’s design is credited to Richard T. Erban, an early 

tractiondrive pioneer, who briefly worked for General Motors during this period (ref. 4). 

In England, after several years of analyzing the Hayes Self-Selector drive, Perbury Engineering, 

Ltd., retrofitted a modified, scrapped Hayes transmission into Hillman Minx sedan in 1958 (ref. 21). 

Fuel savings were reported to be 20 to 25 percent but the concept really never caught on with any of 

the several dozeq companies or so that had previously expressed interest in the drive (ref. 21). 

In the United States in 1959, Charles Kraus installed a modified version of a toroidal CVT into 

an American Motors Nash Rambler (ref. 4). This unit had a semitoroidal roller geometry similar to 

ihat patented in 1932 by Jacob Arter for industrial service. The Arter drive is still commercially 

produced in Switerzerland. In 1962 Mr. Kraus joined CurtisWright Corporation and headed up a 

program supported by the Army Tank-Automotive Center to retrofit this semitoroidal traction drive 

to an Army Jeep. A drive malfunction was encountered after approximately 17 O00 miles of road 

testing (ref. 22). Design improvements were made on a second test drive but this unit also experienced 

operational problems that caused a halt to the program. In 1973 Tracor, Inc., demonstrated a Ford 

Pinto equipped with an improved version of the Kraus drive lubricated with Monsanto’s new traction 

fluid. Although operational characteristics were established, expected fuel economy improvements 

were largely negated by the hydraulic losses in the thrust bearings used to clamp the toroids together 

(ref. 23). More recent toroidal drive designs partially overcame this problem by mounting two 

toroidal drive cavities back-to-back along a common shaft, thereby eliminating these troublesome 

thrust bearings. 

Industrial Traction Drives 

Starting .with the 1877 Hunt drive, adjustable-speed traction drives have been in industrial 

service for more than 100 years. The bulk of these drives has been performing a speed matching 

function for lightduty equipment, such as drill presses. Typical uses are listed in table I. 

A sample of representative tractiondrive configurations appears in figure 8 (ref. 4). According 

to Carson (ref. 24). more than 100 United States patents on adjustable-speed traction drives are on 

file. Out of these, perhaps a dozen or so basic geometries are in production. Some of these drives are 

shown in figure 8. Of those commercially available, few are rated greater than 10-kW power 

capacity. 

125 
TABLE I. - TYPICAL APPLICATIONS FOR TRACTION DRIVES 

Marine propulsion drives 

Earthmoving equipment 

Textile machinery 

Farm equipment, agramachinery 

Rubb‘er mach in ery 

Propel 1 er drives 

Forest products and paper machinery 

Crane drives 

Construction equipment 

Pump drives 

Locomotive and railroad machinery 

Machine tools 

Outdoor tools and recreation vehicles 

Oil field drives and offshore rigs 

Household appliances 

Auto tr an smi s s i on s 

Air conditioning systems 

Steel mill drives 

Mining and ore processing machinery 

The “ball and cone” geometry was commercially introduced in the 1940’s by Jean Kopp in 

Switzerland. The Kopp Variator is said to be the most widely used traction drive, with more than 

250000 units sold around the world through 1975 (ref. 24). A more recent cone-roller variator 

developed by Kopp extends the variator’s power rating to 75 kW in a 582-kg package. 

The wheel and singledisk drive typify the Cartercar and Sears Motor Buggy transmission. 

Because heavy contact loads must be directly reacted by the support bearings, the torque capacity of 

this geometry is greatly restricted. The cone and ring with reducer drive, produced by Graham in 

1935, is one of the earliest traction drives developed in the United States, Graham drives are available 

for ratio ranges to 1O:l and power levels to about 4 kW. The disk-todisk drive, also known as the 

Beier disk drive, was one of the earliest (1949) industrial traction drives capable of handling high 

powers, with current power ratings to 164 kW. The drive components are imported from Japan, 

assembled, and marketed in the United States by Sumitomo Machinery Company. With the possible 

exception of the toroidal geometry, the commercial CVT’s of figure 8 are generally too large and 

complex or have insufficient power capacity to be suitable for advanced automotive applications 

without substantial modifications. An early review of the basic types of adjustable-speed traction 

drives can be found in reference 25. Reference 26 gives descriptive information on 24 types of 

variable speed traction drives that were commercially available in 1%3. 

Early Drive Limitations 

As we have seen, the earliest traction drives generally used leather, wood., rubber, or fiber 

covered friction wheels running against metal disks. As these soft friction materials aged, they lost 

flexibility and wore rapidly. The driving surfaces generally had to be renewed or replaced at frequent 

intervals, depending on the rate of usage. Despite this, friction drives found use in early steam 

tractors, factory machinery, wood-working tools, and’in several vintage cars (ref. 20). These simple, 

smooth, lowcost speed changers are still in use today for lightduty applications ranging from hand 

tools, washing machines, and record players to amusement park rides and cement mixers. In these 

modern drives more durable rubber and reinforced plastic materials have been substituted for the 

leather and fiber wheels of yesteryear. However, the thermal capacity and wear characteristics of 

these softer materials still basically set the useful power capacity of this class of transmission. For 

applications needing a low-cost speed changer, adjustable-speed friction drives are good choices. 

Sizing criteria and typical design information for friction drives can be found in reference 27. 

Oil-lubricated traction drives having hardened steel roller contacts appeared in the early 1920’s. 

Carson (ref. 11) credits Erban in Austria with the-development of the first metal-to-metal, oillubricated drive in 1922. The 1923, Arter drive is another example. The presence of hardened steel 

126 
6 

ORlGlNAC PAGE 

BLACK AND WHITE PHOTOGRAPH 

In t 

t 

c r 

Figure 8 - Typical industrial traction-drive geometries. (Courtesy of Design 

Engineering (ref. 4).) 

rollers in these drives signi.ficantly increased the allowable operating contact stresses. The purpose of 

the oil was to protect the contact surfaces from wear while providing cooling. However, the relatively 

low coefficient of traction of the oil meant that these drives had to carry unusally highcontact loads 

to inhibit slip. High loading generally leads to early pitting, unless the torque rating of the drive was 

appropriately restricted. Even though the early traction drives tended to be bulky, their relatively 

high efficiency and smoothness of operation still made them attractive for many applications. 

The very early friction drives were also vulnerable to slip damage in the event of an overtorque, 

since the normal load to prevent slip was set during assembly and did not change. To overcome this 

problem and to prevent the drive contact from being overloaded during periods of low power, Erban, 

while in Germany, patented in 1921 a-mechanism to automatically regulate the contact normal load 

in porportion to the transmitted torque (ref. 20). Keeping the contact pressures low at light torques 

not only extended fatigue life but helped to reduce unnecessary power loss. Automatic loading 

mechanisms of this type have become regular features of almost all successful traction drives. 

127 
Advancements in Technology 

Traction drive technology made relatively little progress for the first half of this century except 

for the occassional introduction of a new geometric variation. Designs were largely predicated on 

laboratory or field experience and very little of this information was reported in the open technical 

literature. 

Because of the great similarity in the contact operating condition, tractiondrive technology 

benefitted greatly from the wave of technical advancements made for rollingelement bearings. 

Major advancements in bearing design Occhrred in the late 1940’s with Grubin’s work in 

elastohydrodynamic lubrication (ref. 28) and Lundberg-Palmgren’s analysis of rolling-element 

fatigue life (ref. 29). In fact, the lubrication principles, operating conditions, and failure mechanisms 

of tractiondrive contacts and bearing contacts are so similar that the design fundamentals are 

virtually interchangeable. The same may be said for gear contact design criteria as well. 

In view of the durability shortcomings of earlier traction drives, much of the recent research has 

centered on improving the power capacity and reliability of these devices without sacrificing their 

inherent simplicity or high mechanical efficiency. Although work has been performed on many 

fronts, research efforts to date can be loosely categorized under one of several areas: (1) modeling the 

tractive behavior of the lubricant within the contact and its attendant power losses; (2) predicting the 

.useful torque that can be passed between rollers without surface distress or that amount 

corresponding to a given fatigue life; (3) determining and improving the durability characteristics of 

tractiondrive materials, primarily bearing-grade steels; (4) developing lubricants that produce higher 

traction forces in the contact without sacrificing conventional lubricant qualities; and (5) developing 

drive arrangements that maximize durability, torque capacity, and ratio capability and minimize size, 

weight, power loss, and complexity. 

Traetion Phenomena 

A basic understanding of how power is transferred between tractiondrive rollers is helpful in 

reviewing the contribution made in this area. \Figure 9 shows a simple, lubricated, roller pair in 

traction contact. A sufficiently large normal load N is imposed on the rollers to transmit the 

tangential traction force T. The amount of normal load required to transmit a given traction force 

without destructive gross slip is dictated by the available traction coefficient, p, which is the ratio of 

Tto N. Since contact fatigue life is inversely related to the third power of normal load, it is extremely 

desirable to make use of lubricants that produce high values of p. The search for lubricants having 

high traction capabilities will be discussed later. 

The rollers, as illustrated in the enlarged view of the contact appearing in figure 9, are not in 

direct contact but are, in fact, separated by a highly compressed, extremely thin lubricant film. 

Because of the presence of high pressures in the contact, the lubrication process is accompanied by 

some elastic deformation of the contact surface. Accordingly, this process is referred to as 

elastohydrodynamic (EHD) lubrication. Ertel (ref. 30) and later Grubin (ref. 28) were among the first 

to identify this phenomenon, which also occurs for other oil-lubricated, rolling-element machine 

elements such as bearings and gears. The importance of the EHD film in traction contacts lies in its 

ability to reduce and/or eliminate wear while acting as the principal torque transferring medium. An 

excellent accounting of the role of EHD lubrication traction is given by Martin (ref. 31). 

Traction Curve 

The technological properties of the lubricant in the contact, particularly its traction 

characteristics, is fundamental to the design of traction drives. Figure 10 shows a typical tractionversus-slip qurve for a traction fluid. This type of curve is typically generated with a twin-disk 

traction tester (refs. 32 to 35). Imposing a traction force across a lubricated disk contact which is 

rotating at an average surface velocity Ugives rise to a differential surface velocity AU. Three distinct 

regions can generally be identified on a traction curve. In the linear region the traction coefficient 

increases linearly with slip. In the non-Newtonian regions it increases in a nonlinear fashion, reaches 

a maximum, and then begins to decrease. Finally, the curve shows a gradual decay with slip in the 

thermal region due to internal heating within the oil film. It is the linear region of the traction curve 

that is of the greatest interest to traction-drive designers. The design traction coefficient, which 

128 
PRESSURIZED 

~1 T/N 

CS-30-1C1S 

Figure 9. - Power transfer through traction. 

I 

I 

I 

I 

I 

AI I. I 

2 4" M 40 60 m la, 

% SLIP. Ad 

U 

Figure 10. - Typical traction curve showing design range for traction drives. 

dictates how much normal load is needed to transmit a given traction force, is always chosen to be 

less than (by, generally, 20 to 30 percent) the peak available traction coefficient to provide a safety 

margin against slip. Traction drives are generally equipped with a torque-sensitive loading 

mechanism that adjusts the normal contact load in proportion to the transmitted torque. Such 

mechanisms insure that the contact will always have sufficient load to prevent slip without needlessly 

overloading the contact under light loads. 

Traction Coefficient 

Perhaps the single most important factor affecting the torque capacity, life, and size of a 

traction drive is the maximum value of the available traction coefficient. It is this parameter that 

determines the necessary contact load. Since the fatigue life of a contact is inversely proportional to 

the cube of contact load, a 50-percent improvement in the traction coefficient would, for example, 

mean a 240-percent life increase for a given drive under a given set of operating conditions. 

129 
Solidification Behavior 

In a typical tractiondrive contact, severe transient operating conditions are imposed on the 

lubricant. The lubricant is swept into the contact, exposed to contact pressures, which are 10 OOO 

times atmospheric or greater, and,returned to ambient conditions, all in a few milliseconds. Clark, et 

al. (ref. 32). in 1951 were the first to experimentally observe that, under these high pressures and 

shear rates, the lubricant exhibits shearing properties of a plastic solid and will, in fact, yield at some 

critical shear stress. Although it had been known from static high-pressure viscometry experiments 

that an oil’s viscosity would increase dramatically with increasing pressure and eventually solidify 

under high enough pressure, this was really the first time that this effect was observed to occur under 

the highly transient conditions of a traction contact. This solidification phenomenon was later also 

observed by experimentors Smith (ref. 33), Plint (ref. 34), and Johnson and Cameron (ref. 33, 

among others. Indeed, one can see a striking similarity in the shear-stress-shear-strain rate curves of 

a traction fluid and the tensile-stress-strain curves of a typical thermoplastic (fig. 11). It is also clear 

from figure 17 that temperature has a similar adverse effect on the limiting shear strength (maximum 

traction coefficient) of the traction fluid and the yield strength of the thermoplastic. Furthermore, 

both materials exhibit either viscous or elastic behavior depending on the rate of straining. Plastics 

will flow like a liquid if stretched slowly enough but will show elastic resiliancy if pulled suddenly. An 

EHD film in the traction contact shows similar viscoelastic behavior. 

Effects of Operating Conditions 

Whether the lubricant exhibits viscous behavior or acts as an elastic solid depends on a number 

of factors. Contact pressure, temperature, shear rates, and lubricant composition all play important 

roles. Knowledge of the rheological characteristics of the lubricant as it passes through the contact is 

essential to predicting the performance of the contact, particularly in the absence of specific 

experimental data. 

Viscoelastic lubricant behavior has been a subject pursued by a large number of investigators. 

Among them were Crook (ref. 36) and Dyson (ref. 37), who were among the first to call attention to 

such behavior, Johnson and Roberts (ref. 38). who experimentally demonstrated the transition from 

predominantly viscous to predominantly elastic response, and Trachman and Cheng (ref. 39), who 

used a compressional visco-elasticity model to numerical solve for the traction curve. 

Based on the work of these investigators and on the comprehensive traction model advanced by 

Johnson and Tevaarwerk (ref. a), one can assume that for most tractiondrive contacts the 

maximum available traction coefficient is directly related to the shear yield strength of the solidified 

lubricant film. 

From a large body of traction data generated on a twindisk tester (refs. 41 to 44), it was found 

that an increase in contact pressure is beneficial to the available traction coefficient but that an 

increase in surface velocity, temperature, contact ellipticity ratio, misalinement, or spin 

(circumferential slip) have a negative effect. Figure 12 shows the typical effects that pressure and 

speed have on the traction coefficient of a traction fluid from test data generated in reference 45. 

In the linear region of the traction curve, the transfer of torque will cause a small difference in 

velocity to be developed between the surfaces of the driver and driven rollers. This small velocity 

difference, generally less than 1 percent of the rolling velocity, is often referred to as creep rather 

than slip. This is because in creep only part of the contact is experiencing sliding while in slip there is 

total relativebmotion. 

Creep is always present to some extent between rolling bodies that are transmitting torque, 

whether lubricated or not. Carter in 1926 was one of the first to identify the creep occurring between 

a locomotive’s driving wheel and the rail (ref. 46). Typically, a region of microslip will occur between 

the surfaces in the trailing region of the contact while the surfaces in the leading region will be locked 

together without relative motion (ref. 47). As the tangential traction force is increased, the microslip 

region will encompass more and more of the contact until, at some point, the whole contact is in total 

slip. This is the point of impending gross slip. It occurs when the ratio of traction force to normal 

load is equal to the maximum available traction coefficient. When a mechanism to increase the 

130 
SHEAR 

STRESS, 

MPa 

TRACTION FLUID 

120 

m 

TENSILE 

STRESS, 

.MPa SANTOTRAC 50 

u .02 .04 .06 

SHUR STRAIN RATE 

OO 

Figure 11. - Plastic behavior 

MAXIMUM AVAILABLE 

TRACTION COEFFICIENT 

THERM PLASTIC 

6o r WELD STRENGTH/ 

-,' A TEMPERArUrE, 

OC I m23 

THERM PLASTIC 

u 

STRAIN 

OO . 01 .03 .m 

STRAIN 

of a traction fluid. 

MAXIMUM 

CONTACT 

.MPRES SUR E, 

4) 60 m .02 

P CS-RU-ZS~I SURFACE VELOCITY, ds 

Figure 12. - Typical maximum available traction coefficient as function of 

surface velocity and maximum contact pressure. 

traction fluid; a/b = 5; zero spin; temperature, 343 K. Data obtained on a 

twin-disk machine (ref. 45). 

Synthetic hydrocarbon 

normal load with increasing traction is present, the point of impending gross slip is not reached. 

Creep is typically the result of the elastic stretching of the solidified lubricant film and that of the 

roller material. The lubricant film gets stiffer with increasing pressure and because it is so thin 

relatively little strain occurs. In fact, as pointed out by Tevaarwerk (ref. 48), for many traction drives 

that regularly operate at contact pressures above 1.2 GPa, most of the creep actually takes place in 

the steel rollers and not in the EHD film. Johnson, et al. (ref. 49), have developed relatively simple 

methods to determine how much of the elastic effects ;ire due to the roller material and how much are 

due to the film. 

The amount of creep in a lubricated tractiondrive contact can vary from as little as 0.1 percent 

of the rolling velocity at high pressures and low speeds to 3 or 4 percent or more when the pressures 

are low, speeds are high, and sideslip (misalinement) or spin (circumferential contact slip) is 

appreciable. Although these creep values are quite small relative to gears, where sliding velocities can 

be 30 or 40 percent of the pitch-line velocity at the tooth engagement and disengagement points, it is 

newrtheless important to minimize the creep value through the proper selection of operating 

conditions. geometry. and lubricant. Every percentage point loss in creep represents a percentage 

poirir l~w in iped arid n corresponding percentage point loss in mechanical efficiency. 

131 
Traction Fluids 

Because of the importance that the coefficient of traction has on the life, size, and performance 

of a traction drive, considerable attention has been given to identifying fluids with high traction 

properties, starting in the late 50’s with Lane’s experiments (ref. 50). Using a modified two-ball 

tester, Lane found an apparent inverse relationship between the traction coefficient and temperatureviscosity index for naphthenic and paraffinic mineral oils. 

In a comprehensive investigation into traction phenomena, Hewko (ref. 42) obtained traction 

performance data for both compounded and Uncompounded mineral oils as well as for a group of 

synthetic fluids. He also investigated the effect of oil additives on traction as well as the effects of 

operating conditions, roller geometry, and surface topography. His results indicated that the 

lubricant composition and surface topography had the greatest overall effects on traction and that 

naphthenic-based mineral oils gave better performance than paraffinic oils. He also observed that 

many common types of additives can markedly reduce traction (ref. 42). 

The research of reference 51 describes the development of a formulated traction fluid, 

designated as Sunoco Traction Drive Fluid-86. This fluid was subsequently field tested in the General 

Motors turbine bus toric drive and in an automatic transmission for the Oldsmobile Toronado with 

reportedly good results. 

Hammann, et al. (ref. 52), examined some 26 test fluids and, unlike Lane (ref. 50), could not 

find any obvious relationship between the coefficient of traction and the viscosity index. However, 

his tests did identify several synthetic fluids that had up to 50 percent higher coefficients of traction, 

depending on test conditions, than those reported for the best naphthenic base oils (ref. 52). This 

research laid the ground work for the development of Monsanto’s family of commercial traction 

fluids, Santotrac 30, 40, 50. and 70. These fluids are the most widely used traction oils today. The 

results of accelerated five-ball fatigue tests (ref. 53) indicated that these synthetic cycloaliphatic 

traction fluids have good fatigue-life performance, comparable with the reference tetraester oil used 

in this experiment. 

In an unusually complete doctorial investigation into the nature of traction power transfer, 

Gaggermeier (ref. 44), in one part of his work, measured the coefficient of traction of 17 lubricants 

on a twin-disk traction tester at both high and low contact pressures and surface speeds. The traction 

fluids in his tests showed substantially higher coefficients of traction than all of the commercial 

naphthenic mineral oils tested. The greatest differences occur at relatively low pressures and high 

surface speeds (fig. 13). At relatively high pressures and low speeds the traction fluids show less of an 

advantage. Under such conditions a good quality naphthenic mineral oil would serve almost as well. 

However, for most tractiondrive applications there is considerable incentive to using a traction fluid, 

with expected traction improvements falling somewhere between the two examples of figure 13. 

What is noteworthy about the traction curves in figure 13 is the way they illustrate the common 

and important tendency of traction fluids to solidify into a plastic material at far lower pressures and 

high speeds than conventional oils (e+, see the work of Blair and Winer, ref. 54). This is the basic 

reason why the traction coefficient for the two traction fluids are already at high levels in figure 13(a), while those for the mineral oils need substantially higher pressure to attain high levels as 

indicated in figure 13(b). The tendency for early solidification has a lot to do with what makes a 

traction fluid a traction fluid. 

I 

Capacity and Durability 

Failure Criteria 

Traction .drives, like rolling-element bearings are generally sized on the basis of rolling-element 

fatigue life. This is because, for most applications, other than those that are particularly short-lived, 

the stress levels required for acceptable fatigue life are generally well below those for static yield 

failure. Because of these relatively low maximum operating stress levels, traction drives can generally 

tolerate shock loads several times the maximum design value without plastic deformation or other ill 

effects. This is contrary to the misbelief that traction drives are particularly vulnerable to sudden 

overloads. Funhermore, if these transient overloads are brief and do not occur too frequently, only a 

relatively small penalty to the drive’s total fatigue life will result. 

132 
TRACTION 

COEFFICIENT, 

(r 

LOW PRESSURE, HIGH SPEED 

MAXIMUM HERTZ PRESSURE 1.09 CPa 

SURFACE VELOCITY 8.3 rnlsec 

OIL IN TEMPERATURE WC 

r SANTOTRAC 70 

. 

TRACTION FLUID .an 

. uc 

OO 

u 

.02 .04 .06 

HIGH PRESSURE, LOW SPEED 

MAXIMUM HERTZ PRESSURE = L 86 

SURFACE VELOCITY = 0.83 rnlsec 

OIL TEMPERATURE = 5@C 

rSANTOTRAC 70 *---- 

LSHEU CLAVUS 

VAC HLP 25 

I I 

I .02 .04 .06 

' GPa 

CREEP RATE. AUU 

Figure 13. - Traction characteristics of a traction fluid compared with 

mineral oils (ref. 44). 

A traction drive's sensitivity to shock loads is also dependent on the ability of the contact surface 

to avoid skidding or heating damage. If the drive is equipped with a fast-acting loading mechanism, 

skidding is unlikely. Also the transient nature of shock loads is such that there is usually insufficient 

time to overheat the contact, thus avoiding surface damage. However, sustained overloads can cause 

a thermal brewown of the EHD film leading to failure. 

Fatigue Life 

The normally expected failure mode of a properly designed traction drive will be rollingelement 

fatigue. This failure criterion is exactly analogous to pitting failure in gears and spalling failure in 

rollingelement bearings. The risk of wear or scuffing failures of traction drive contacts can be 

eliminated or greatly minimized through the use of proper materials and also proper lubricating and 

cooling design practices such as those that have been successfully applied in bearing and gear design. 

In view of this similarity in the failure mechanism, rollingelement bearing fatigue-life theory can be 

used to determine the expected service life of a traction drive. 

In 1947 Lundberg and Palmgren (ref. 29) published a statiscal theory for the failure distribution 

of ball and roller bearings. They theorized that subsurface-originated fatigue pitting was due to high 

stresses developed at a subsurface defect in the bearing material. The probability of failure was 

related to the value of the maximum orthogonial reversing shear stress, the depth below the surface 

where this stress occurs and the volume of material that l's being stressed. This is the theory on which 

bearing manufacturers, still today, base ball and roller bearing service-life ratings. 

In 1971 an ASME life adjustment factors guide (ref. 55) was published to adjust the life ratings 

predicted by Lundberg and Palmgren analysis for recent advancements in bearing design, materials, 

processing, and mmfacturing techniques. 

In 1976 Coy, et ai. (ref. 56), published a contact fatigue-life analysis for traction drives based on 

Lundberg-Palmgren theory. This analysis was used to predict the service life of a toroidal traction 

drive. In a,1980 publication, Rohn, et al. (ref. 57), presented a simplified version of this fatigue-life 

analysis and used it to show the effects of torque, size, speed, contact shape, traction coefficient, and 

number of multiple contacts on the predicted drive life. Their results show that multiple, load-sharing 

arrangements significantly benefit torque capacity and drive life, and that fatigue life is proportional 

to the 8.4 power of size for constant torque and traction coefficient. 

The aforementioned Lundberg-Palmgren fatigue life analyses can be used with reasonable 

certainty to determine the durability characteristics of most traction drives. However, not presently 

considered in the analyses is a means to account for the potentially adverse effects of traction. Some 

133 
investigators (Vf. 58) have found up to a several-fold decrease in rolling-element fatigue life when 

relative sliding and traction have been introduced; while others (ref. 59) have observed no change or 

even an improvement in fatigue life, depending on the direction of the applied traction force. A new 

analytical fatigue-life model, proposed by Tallian, et al. (ref. 60), considers the effects of surface 

traction on surface and subsurface-originated spalling. This work is significant because it addresses 

surface spalling, the more likely fatigue failure mode of gear and traction drive contacts. However, in 

its present form, the model requires many metallurgical and surface microgeometry parameters, 

which are not readily available. Thus, the utility of the analysis is, at present, limited. Also, there are 

currently insufficient test data to properly substantiate this model or to develop universal fatigue-life 

adjustment factors for the possible negative effects of surface traction. An element of conservatism is 

in order when establishing service life ratings based on current prediction methods. 

Materials 

Earlier traction drives were not exploited to their full potential because of uncertainties 

regarding their longterm reliability. The limited durability characteristics of the materials used in 

these drives was a major contributing factor. The substitution of oil-lubricated, hardened steel roller 

components in place of rubber or reinforced plastic running dry against cast-iron parts raises their 

load capacity by at least an order of magnitude (ref. 27). 

Because of the similarity in operating conditions, hardened bearing steels are logical choices for 

tractiondrive rollers. Today’s bearing steels are of significantly higher quality than the traditional 

air-melted, AISI 52100 steels used in rolling-element bearings since the 1920’s. The introduction of 

vacuum remelting processes in the late 1950’s has resulted in more homogeneous steels with fewer 

impurities and have extended rolling-element bearing life several-fold (ref. 61). Life improvements of 

eight times or more are not uncommon according to reference 55. This reference recommends that a 

life-improvement factor of six be applied to Lundberg-Palmgren bearing life calculations when using 

modem vacuum-melted AISI 52100 steel. A similar life-improvement factor is applicable to tractiondrive life calculations. This improvement in steel quality in combination with improvements in 

lubricant traction performance have increased the torque capacity of traction drives several-fold. 

Performance Predictions 

Contact Traction 

The distribution of local traction forces in the contact of an actual traction drive can be rather 

complicated as illustrated in figure 14. This figure shows the distribution of local traction vectors in. 

the contact when longitudinal traction, misalinement, and spin are present. These traction forces will 

dine themselves with the local slip velocities. In tractiondrive contacts some combination of 

traction, misalinement, and spin are always present. To determine the performance of a tractiondrive contact, the elemental traction forces must be integrated over the contact area. 

Because of the parabolic pressure distribution, the elemental traction forces are largest near the 

center of the contact and diminish in magnitude near the contact perimeter. As expected, in the case 

of longitudinal traction (fig. 14(a)), the forces aline themselves in the rolling direction. With the 

addition of misalinement (fig. 14(b)), a sideslip velocity is introduced causing the vectors to cock in 

line with the sideslip angle. Using conical rollers generally results in a circumferential slip pattern 

referred to as spin (fig. 14(c)). This rotary motion is due to the fact that the contact is in pure rolling 

only at its center. At the right hand edge of the contact the upper roller is sliding over the lower roller 

because of the mismatch in contact radii. At the left hand edge the situation is reversed and so is the 

direction ofislip. Because of solid-body rotation, a complete pattern of spin exists over all of the 

contact. 

The power throughout the contact is determined from a summation of the traction force 

components dined in the rolling direction times their respective rolling velocities. It is clear that in 

misalinement only a portion of the traction force is generating useful traction and that the remainder 

is generating useless side force. For pure spin no useful traction is developed, since the elemental 

traction forces cancel one another. Since the contact power loss is proportional to the product of the 

elemental traction forces and slip velocities, the presence of spinand misalinement can significantly 

134 
TRACTION 

f ROCLING 

DIRECTION 

TRACTION AND 

MISALIGNMENT 

ANGLE \I 

\ 

A t 

Figure 14. - Effect of misalinement and spin on contact traction force vectors. 

decrease the efficiency of the contact. Furthermore, both conditions lower the available traction 

coefficient and reduce the amount of torque that can be transmitted safely. Designs that minimize 

spin and side slip can be quite efficient. Contact efficiency of 99 percent or higher are possible. 

Traction Experiments 

Although there have been previous attempts to analyze the losses associated with ball bearings 

(ref. 62). the first systematic attempt to research the losses of tractiondrive contacts occurred at the 

“Institut Fur Maschinenelemente und Forder Technik” located at the Technical University of 

Braunschweig in German): between 1955 and 1960 (refs. 63 to 67). These investigations, which are 

summarized in an excellent manner by Wernitz (ref. 68), include analysis and test work on both 

simple traction contact test machines and on a commercially available Kopp ball variator. Tables and 

plots were developed which permit the calculation of friction losses due to creep (microslip) and spin 

(circumferential slip). This analysis treats the EHD film in the contact as behaving as perfectly 

plastic, that is, the material yields at some critical shear stress, exhibiting no elastic behavior. This 

model gives satisfactory results for contacts experiencing relatively high local strain rates such as 

those drives that have moderately high spin. In reference 69 Magi takes a similar, but more general 

approach and gives examples with experimental justification. 

In the United States some of the earliest investigations into traction contact phenomena as they 

relate to traction drives were conducted by Hewko (refs. 42, 43, and 70). As mentioned earlier, 

Hewko (ref. 42) obtained basic traction performance data for various roller geometries, operating 

conditions and lubricants. He later extended much of these data to roller contacts that operate at very 

high surface speeds (up to 127 m/sec) (ref. 23). Much of this test information served as a data base 

for the construction of several fixed-ratio, simple planetary traction drives, of the type used in 

English life boats back in the 1930’s (ref. 71). One of these units, a 6.25 to 1 speed reducer attached to 

a 4.5-kW electric motor, was used as the main drive for General Motors oceanographic submarine. A 

second, 6 to‘l reduction ratio drive delivering 373 kW at SO00 rpm was built for a turbine-powered 

torpedo system. In reference 70 Hewko compared the efficiency and noise characteristics of a 75-kW, 

3.5 to 1 ratio, planetary tracton drive with that of a class 8, planetary gear drive of similar size, ratio, 

and power capacity. His test results showed that the traction drive had not only better part-load 

efficiency but provided substantial noise reductions as well. An overall sound pressure level of just 

70 dB, about conversation noise level, was recorded for the traction drive at a particular operating 

condition. By contrast, at this same condition the gear drive registered 94 dB, or 16 times the noise of 

135 
the traction drive. The set of noise signature traces from this experiment with the integrated overall 

sound pressure levels appear in figure 15 (ref. 70). The quiet running characteristics of traction drives 

are undoubtedly appreciated by those who design drive systems for passenger carrying vehicles or for 

factory equipment under current OSHA noise standards. 

Theory 

In the 1960’s numerous papers were presented on the prediction of traction in EHD contacts. 

Some of these included Cheng and Sternlicht (ref. 72), Bell, Kannel, and Allen (ref. 73), and Dowson 

and Whitaker (ref. 74). In the early 1970’s simplifications and refinements in the traction models 

took place as exemplified by Dyson (ref. 37)’ Kannel and Walowit (ref. 73, and Trachman and 

Cheng (ref. 39). 

About this time, Poon (ref. 76) and Lingard (ref. 77) developed grid methods to predict the 

available traction forces of a contact experiencing spin. Poon’s method utilized the basic traction 

data from a twindisk machine together with contact kinematics to predict the available traction. 

Lingard used a theoretical approach in which the EHD film exhibited a Newtonian viscous behavior 

at low shear rates until a critical limiting shear stress was reached. At this point the film yielded 

plastically with increasing shear rate. This model showed good correlation with experimental traction 

data from a toroidal, variable-ratio drive of the Perbury type. This same model was also used 

successfully by Gaggermeier (ref. 44) in an comprehensive investigation of the losses and 

characteristics of tractiondrive contacts. In addition to copious amounts of twindisk traction data 

for numerous lubricants under various combinations of slip, sideslip, and spin, Gaggermeier 

(ref. 44) also investigated the sources of power losses of an Arter type toroidal drive. His findings 

were that, of the total power losses, the loaddependent bearing and drive idling (no-load) losses were 

always greater than the losses due to traction power transfer. This underscores the need to pay close 

attention to these tare losses in order to end up with a highly efficient traction drive. 

The most recent and perhaps most comprehensive traction contact model is that proposed by 

Johnson and Tevaarwerk (ref. 40). Their model covers the full range of viscous, elastic, and plastic 

behavior of the EHD film. The type of behavior depends on the Deborah number, a relative measure 

of elastic to inelastic response of the lubricant film, and the strain rate. At low pressures and speeds 

(low Deborah number), the film exhibits linear viscous behavior at low strain rates, becoming 

increasingly more nonlinear with increasing strain rate. At higher pressures and speeds, more typical 

of tractimdrive contacts, the response is linear and elastic at low rates of strain. At sufficiently high 

strain rates, the shear stress reaches some limiting value and the film shears plastically as in the case, 

of some of the earlier traction analytical models. 

SOUND 

PRESSURE 

LEVEL, 

DB 

I GEAR DRIVE 

TRACTI3N DRlM 

4- CELL BACKGROUND NOISE 

FREQUENCY - CPS 

Figure 15. - External noise spectra for traction drive compared with planetary 

gear reducer at 2000 rpm input speed and- 136 N-m input torque. (ref. ‘70). 

136 
In reference 48 Tevaarwerk presents graphical solutions developed from the Johnson and 

Tevaarwerk elastic-plastic traction model. These solutions are of practical value in the design and 

optimization of tractiondrive contacts. By knowing the initial slope (shear modulus) and the 

maximum traction coefficient (limiting shear stress) from a zero-spinhero-sideslip traction curve, the 

traction, creep, spin torque, and contact power loss can be found over a wide range of spin values 

and contact geometries. 

Figure 16 shows that this analysis compares favorably with test data, taken from experiments of 

Gaggermeier (ref. 44). The observed, pronounced reduction in the available traction coefficient with 

just a few degrees of misalinement, underscores the need to maintain accurate alinement of roller 

components in traction drives. 

The Johnson and Tevaarwerk model, like many of the previous theories, provides an isothermal 

solution to the ascending portion of the traction curve. Under most circumstances the results from 

this isothermal analysis are quite satisfactory from a design standpoint. However, at higher surface 

speeds and for substantial spin, thermal effects start to become important. In reference 78 Daniels 

introduces a temperaturedependent elastic shear modulus term into the model, but this seems to 

have a rather weak effect on the results. Using an elaborate heat balance computer analysis, 

Tevaarwerk (ref. 79) in 1979 was able to correctly predict the heating effects occumng in a traction 

contact; but his solution is a bit too formidable for general design purposes. In reference 80 

Tevaarwerk presents a simple, empirical technique to predict the thermally influenced, large-spin, 

traction curve from the large slip portion of a no-spin traction curve, but this requires knowledge of 

the full curve, which is not always available. Simpler analytical techniques are now being investigated 

(refs. 81 and 82) to better predict how heat generation within the contact affects the available traction 

coefficient. 

Fluid Traction Data 

To be able to apply the aforementioned traction drive, certain fundamental fluid properties, 

namely, the lubricant's shear modulus and limiting yield shear stress, must first be known under the 

required operating speeds, pressures, and temperatures. Because of the difficulty of simulating the 

highly transient nature of an actual traction contact, the most reliable basic fluid property data are 

deduced, using methods described in reference 44, from the initial slope and maximum traction 

coefficient of an experimental traction curve. Unfortunately, there has been a general scarcity of 

these types of data over sufficiently broad enough operating conditions for design purposes. 

Recently, experimental traction data were obtained under a NASA program for both the Monsanto 

and Sun Oil traction fluids over a range of speeds, pressures, temperatures, spin, and sideslip values 

that might be encountered in traction drives (ref. 45). A regression analysis applied to the data 

resulted in a correlation equation that can be.used to predict the initial slope and maximum traction. 

coefficient at any intermediate operating condition (ref. 45). 

Recent Developments 

During the past 5 years, several traction drives, which incorporate much of the latest technology, 

have reached the prototype stage. Laboratory tests and design analysis of these drives show them to 

have relatively high-power densities and, in some cases, to be ready for commercialization. 

Nasvytrac Drive 

Although lightduty variable-ratio traction drives have been reasonably successful from a 

commercial standpoint, very few, if any, fixed-ratio types have progressed past the prototype stage. 

This is somewhat surprising in view of the outstanding ability of traction drives to provide smooth, 

quiet power transfer at extremely high or low speeds with good efficiency. They seem particularly 

well suited for high-speed machine tools, pump drives, and other turbomachinery. In other industrial 

applications they offer potential cost advantages because traction rollers should not be much more 

expensive to manufacture in quantity than ordinary rollers in roller bearings. 

In terms of earlier work on fixed-ratio traction drives, the developmental effort at General 

Motors Research Laboratories on their planetary traction drive (as described by Hewko (ref. 70)), 

137 
TRACTION 

COEFFICIENT 

v 

CAGGERMEIER J & T 

DATA ANALYSIS SANTOTRAC 50 

k = 1.6 

a 

A 

0 

k 

- 

= 1.0 MAX PHz = 1.8 GPa 

-- $; =-E kc --- 

EFFECT OF SIDE SLIP ON V 

B, 

EFFECT OF SPIN ON v 

&g 

* IJ --.9 

.lou #I -----a 

//-- 2a 

/- 

*&It 

5lM SLIP 8 

+ 

I 

0 .02 .M .06 .08 0 .02 .04 .M 

SLIP, AulU 

Figure 16. - Comparison of Johnson and Tevaarwerk analysis (ref. 48) with 

Gaggermeier test data (ref. 44). 

was perhaps the most complete. As mentioned earlier, several of these drives were built and tested, 

including a 6-to-1 ratio, 373-kW unit for a torpedo and a 3.5-to-1 ratio, 75-kW test drive. This last 

drive exhibited better efficiency and lower noise then a comparable planetary gear set. 

Interest in fixed-ratio traction drives is also high outside of the United States. Tests were recently 

conducted in Japan on a planetary traction drive of a construction similar to the General Motors unit 

for use with a gas-turbine awrilliary propulsion unit (APU) system (ref. 83). Planetary traction drives 

have also been studied in Finland. 

The traction drives described thus far have a simple, single-row planet-roller format. For drives 

like these the number of load sharing planets is inversely related to the speed ratio. For example, a 

four-planet drive would have a maximum speed ratio of 6.8 before the planets interfered. A fiveplanet drive would be limited to a ratio of 4.8 and so on. 

A remedy to the speed ratio and planet number limitations of simple, single-row planetary 

systems was devised by Nasvytis (ref. 84). His drive system used the sun and ring-roller of the simple 

planetary traction drive, but replaced the single row of equal diameter planet-rollers with two or 

more rows of “stepped” or dualdiameter planets. With this new “multiroller” arrangement, 

practical speed ratios of 250 to 1 could be obtained in a single stage with three planet rows. 

Furthermore, the number of planets carrying the load in parallel could be greatly increased for a 

given ratio. This resulted in’ a significant reduction in individual roller contact loading with a 

corresponding improvement in torque capacity and fatigue life. 

In reference 84 Nasvytis reports the test results for several versions of his multiroller drive. The 

first drive tested was a 373-kW (500-hp) torpedo drive of three-planet row construction-with a 

reduction ratio of 48.2 and an input speed of 53 OOO rpm. The outside diameter of the drive itself was 

43 cm (17 in.), and it weighed just 930 N (210 lb) including its lightweight magnesium housing. It 

demonstrated a mechanical efficiency above 95 percent. To investigate ultrahigh-speed operation, 

Nasvytis tested a 3.7-kW (5-hp), three-row, 120-to-I ratio speed increaser. The drive was preloaded 

and operated without torque at 480 OOO rpm for 15 min and ran for 43 consecutive hr at 360 OOO rpm 

without lubrication but with air cooling. Two back-to-back drives were operated for 180 hr at speeds 

varying from lo00 to 120 OOO rpm and back to lo00 rpm. They transmitted between 1.5 and 2.2 kW 

(2 and 3 hp). Another 3.7-kW (5-hp), three-row speed increaser, with a speed ratio of 50, was tested 

for more than 5 hr at the full rated speed of 150 OOO rpm with oil-mist.lubrication and air cooling. It 

successfully transmitted 3.7 kW (5 hp) at 86 percent efficiency (ref. 84). 

138 
The basic geometry of the Nasvytis traction (Nasyvtrac) drive is shown in figure 17. Two rows of 

stepped planet-rollers are contained between the concentric, high-speed sun roller and low-speed ring 

rollers. In the drive shown the planet rollers do not orbit but are grounded to the case through 

relatively low-speed and lightly loaded reaction bearings that are contained in the outer planet row 

only. The high-speed sun-roller and other planet bearings have been eliminated. The sun-roller and 

first row rollers float freely in three-point contact with adjacent rollers for location. Because of this 

floating roller construction, an excellent force balance situation exists even with thermal or 

mechanical housing distortion or with slight mismatches in roller dimensions. 

Based on the inherent qualities of the Nasvytrac drive, a NASA program was initiated (refs. 85 

and 86) to parametrically test two versions of the drive. These drives of nominally 1440-1 ratio were 

tested at speeds to 73 OOO rpm and power levels to 180 kW. Parametric tests were also conducted with 

the Nasvytis drive retrofitted to an automotive gas-turbine engine. The drives exhibited good 

performance, with a nominal peak efficiency of 94 to 96 percent and a maximum speed loss due to 

creep of approximately 3.5 percent. The drive package size of approximately 25 cm diameter by 

11 cm wide (excluding shafting) and total weight of about 26 kg (58 Ib) makes the Nasvytrac drive, 

with a rated mean life of about 12 OOO hr at 75 kW and 75 OOO rpm, size competitive with the best 

commercial gear drive systems. 

A 70 OOO rpm, 10.8-to-1-reduction-ratio Nasvytrac drive weighing just 4 kg was designed and 

built for a long-life, rocket-engine pump drive system to drive the low-speed liquid-oxygen and 

liquid-hydrogen boost pumps (ref. 87). Either an auxilliary turbine or a gear drive off the main pump 

can be used. Use of an auxilliary turbine complicates the design, while gear drives were not well suited 

for this application because they wore badly in a matter of 20 min or so in this hostile cryogenic 

environment. This fell far short of the 10-hr life requirement envisioned for future, reuseable rocket 

engines. The relatively low sliding characteristics of the Nasvytrac drive, coupled with its 

demonstrated ability to run for long periods of time unlubricated, make it an excellent candidate for 

this application. Preliminary tests on this drive in liquid oxygen, including tests in which the drive 

.was repeatedly accelerated under full power (15 kW) to 70 OOO rpm in 5-sec intervals, showed it to 

perform satisfactorily (ref. 87). Cumulative operating times up to an hour have been recorded. 

Future work is needed to realize the 10-hr life goal, but the potential of this transmission for this 

application has been clearly demonstrated. 

Current work with variations of the Nasvytis tractiondrive concept are underway at NASA. A 

recently fabricated, 370-kW (500-hp) helicopter main rotor transmission combines the best features 

of gears with traction rollers. This experimental hybrid transmission, which offers potential cost, 

noise, and reliability benefits, will be tested shortly. Test of a 2240-kW (3000-hp) hybrid, twinengine 

helicopter transmission is also planned. This 8 140-1 -reduction-ratio, traction transmission will be the 

most powerful ever tested, carrying an output torque in excess of 73 OOO N-m (650 OOO inslb) in an 

estimated 515-kg package. Hybrid and pure-traction Nasvytis drives are also being considered for 

wind-turbine applications where a low-cost but highly reliable speed increaser is needed to drive the 

high-speed alternator. Also,-tests are being conducted on an infinitely variable-ratio Nasvytis drive, 

but the performance data are still too preliminary to be reported here. 

. 

Promising Variable Speed Drives 

Taking advantage of the latest technology, several designers have attempted to develop traction 

CVT’s for automotive use. The potential of improving the city fuel mileage 20 to 25 percent, or more, 

of cars normally equipped with three- or four-speed automatic transmissions (ref. 88) or of doubling 

the fuel mileage in the case of flywheel equipped cars (ref. 89) has been the major incentive for the 

resurgence in automotive CVT research and development. These applications represents a significant 

challenge, since compactness, efficiency, cost, and reliability are all at a premium. 

Perbury CYT. -One such automotive effort is that being conducted by BL Technology, Ltd., 

formerly British Leyland, on a Perbury, double-cavity toroidal drive. This concept is rather old (first 

patented in the U.S. by C. W. Hunt in 1877). It is also rather well-explored, as mentioned earlier, 

having been investigated by the General Motors Research Laboratory in the early 30’s and late ~O’S, 

demonstrated in a 1934 Austin-Hayes, later in a 1957 Hillman-Minx, and also in a 1973 Ford Pinto, 

but with offset rollers. 

139 
SUN WR7 FIRST ROW OF 

PlANET ROCLERS-, ,’ ,-SECOND ROW OF 

MECHANISM J BEARINGS 

Figure 17. - Geometry of the Nasvytis traction (Nasvytrac) test drive 

(ref. 85). 

In 1977 Lucas Aerospace in England adapted the Perbury-drive for maintaining constant 

frequency of the AC generators on the Sidley hawker Harrier, a VSTOL jet fighter. This single 

cavity, toroidal drive is suitable for driving aircraft generators having output ratings up to 30 kV A. 

More than 20 years before this, Avco Lycoming, in the U.S., also offered a line of mechanical 

constant-speed drives based on the toroidal tractiondrive principle. They were used on several 

aircraft in the ~O’S, including the Douglas A-4E fighter, but have long since been discontinued. 

In the case of the Leyland-Perbury automotive CVT (fig. 18; ref. 90), the double-toroidal drive 

cavities have six tiltable transfer rollers between input and output toroids. A hydraulically controlled 

linkage system can tilt these rollers from one extreme position to another. By combining this toroidal 

drive with a two-range, output planetary gear system, the overall transmission ratio range is greatly 

expanded. 

The BLIPerbury transmission was installed in a medium size test car having a fourcylinder, 

60-kW engine. The test car showed fuel mileage improvement of 15 to 20 percent for an average mix 

of European driving (ref. 90). Also, acceleration times were comparable with a manual transmission 

car having 10 percent higher power to weight ratio and driven by a skilled driver. However, the future 

production picture for this transmission is not clear. 

Yudetec CYT. -A promising traction CVT that is of a rather new vintage is the nutating drive 

being developed by Vadetec Corp. As shown in figure 19 (refs. 91 and 92) a doubleconical-roller 

assembly, complete with an automatic loading mechanism, is mounted at an angle in a drive cylinder 

that is driven by the input shaft. As the input shaft rotates, the doublecones perform a nutating 

motion and at the same time are forced to rotate about their own axis as they make drive contact with 

a pair of moveable control rings. These rings are grounded to the housing but can be axially moved 

together or apart. A gear pinion attached to the end of the cone shaft orbits the output shaft axis at 

input shaft speed, while spinning about its own axis, due to cone rotation, in the opposite direction. 

By varying the axial position of the control rings, the rolling radius of the cones can be synchronously 

changed. This, in turn, causes a change in rotational speed of the cone shaft pinion but does not 

affect the pinion’s rotating speed, which occurs at input shaft frequency. 

Since the cone shaft pinion is in driving engagement with the ouput shaft through one of several 

possible interchangeable gear arrangements, the variation in pinion rotational speed causes a 

140 ORIGINAL PAGE 

BLACK AND WHITE PHOTOGRAPH 
+ 

ORlGlNAC PAGE 

BLACK AND WHITE PHOTOGRAPH 

CS-80-4879 d 

Figure 18. - BL TechnologyIPerbury traction CVT, 60-kW passenger car test 

installation (ref. 90). 

INPU 

SHAF 

.E 

RINGS 

Figure 19. - Vadetec nutating traction CVT. (Courtesy of Vadetec Corp., Troy, 

corresponding variation in output shaft speed. By changing the output gearing, the pinion speed can 

either vectorially subtract from pinion/input shaft orbital speed, allowing the ouput shaft speed to 

reach zero, if desired, or vectorially add to input orbital speed. The latter arrangement results in a 

transmission with only overdrive capability. 

A couple of Vadetec CVT's have already been built and tested. One of these prototypes has 

shown successful operation as part of a tractor drive train. Although this transmission is still in the 

development and acceptance stage, it is a good example of the new breed of traction drive. 

Michigan.) 

141 
Fafnir CYT. -A planetary, cone roller type CVT under development by the Fafnir Bearing 

Division of Textron appears in figure 20 (ref. 93). This transmission is directed toward the mobile 

equipment market, particularly as a replacement to hydrostatic transmission in garden and lightduty 

tractors up to about 37 kW (50 hp). A set of double-sided, conical traction rollers are trapped 

between a pair of inner races and a pair of outer races. The conical rollers are mounted in a carrier 

that drives the planet gears in the output gear differential. The gear differential serves the purpose of 

expanding the limited ratio range of the traction drive to cover output speeds from forward to 

reverse, if desired. This is an example of power-recirculating transmission. The traction inner races 

are splined to the sun gear shaft of the planetary which is, in turn, keyed to the input shaft. To change 

the ratio, the outer traction race halves, which are grounded to the housing, can be either manually or 

hydraulically, in a later version, pushed together or spread apart. This causes the cone rollers’ rolling 

radii at the outer-race contact to increase or decrease while simultaneously causing the opposite to 

occur at the inner-race contact. The change in rolling radii causes a corresponding change in the orbit 

speed of the cones and, thus, output-shaft speed. Because the inner-race load-springs, the 

transmission automatically “down shifts” with increased load. The current design was specifically 

intended to be one of low cost with maximum flexibility, so a single range output planetary was 

selected. Consequently, a high degree of power recirculation exists, limiting efficiency to about 

85 percent maximum for the current design (ref. 6). A more efficient version would be needed for 

automotive applications. 

Other traction CVT’s. -The traction developments previously cited are by no means the only 

efforts underway. Many of the industrial manufacturers are no doubt upgrading the ratings of their 

systems by taking advantage of the latest technology. Fwthermore, variations of older concepts are 

being reevaluated and upgraded with design improvements. Two notable examples are the 

AiResearch toroidal drive and the Bales-McCoin cone-roller CVT. Both of these systems were carried 

to the preliminary design stage under a NASA contract for DOE to develop CVT’s for use in 

advanced electric vehicles. 

The AiResearch design (ref. 94) is a doublecavity toroidal CVT, containing two power rollers 

per cavity. The double toroidal drive elements, similar in many respects to the toroidal drive units 

mentioned earlier, are permanently connected to differential gearing. 

The Bales-McCoin unit (ref. 95) consist of a central input traction roller surrounded by four 

cone-rollers which in turn are connected to the output planetary differential via idler gears. The cones 

are hydraulically loaded against the central roller using a novel microprocessor control system. Based 

on current slip rate signals, the control system adjusts the normal load between rollers to the 

minimum required to prevent significant roller slip at any given oprating condition, 

Both of the above CVT’s are in early stages of development. Estimated weights and sizes of these 

transmissions are comparable with equivalent conventional automatic transmissions, and calculated 

efficiencies are generally in the low 90’s. General information on these CVT’s can be found in 

reference 96. 

t=p RATIO AOJUSTCR 

TRACTION MIIVL noun INNER RACES 

RING GEAR 

LOAD SPIIINGS 

PLANET GEAR 

outm RACES 

uu 

Figure 20. - Fafnir Bearing Division’s planetary sone roller traction drive (ref. 93). 

142 
Future Technology Requirements 

t 

L 

Although there has been significant progress within the last decade in the development of 

technology for traction drives, future refinements in certain areas are still needed. 

Traction Fluids 

Today’s traction fluids are far superior to the earlier lubricants used in traction drives. However, 

modem traction fluids generally have poor viscosity index characteristics. At temperatures well 

below freezing, traction fluids can be too viscous to pour properly. Pour-point suppressents such as 

those incorporated in conventional oils for Arctic service might help. 

Traction fluids and the additives that they contain have a tendency to capture and retain air at 

temperatures below about 70” C. Aeration of the oil is undesirable from both a cooling and hydraulic 

control system standpoint. This problem is minimized when the oil is subjected to some nominal 

pressure which tends to significantly condense trapped air bubbles, as occurs in the contact between 

rollers. 

The key to faster and more powerful traction drives is the development of traction fluids whose 

coefficient of traction shows less degradation with higher surface speeds, operating temperatures, 

and contact spin. Higher traction at lower contact pressures is also desirable for drives that have extra 

long life requirements. Today’s synthetic traction fluids appear to be reasonably stable but further 

work may be needed for “sealed-for-life” drive systems. 

. Materials 

Although today’s bearing steels are well suited for traction drives, durability improvements and 

lower material costs are always welcome. Studies are required to determine if special coatings or 

* nonmetallic materials might be useful in certain applications. 

Sizing and Design Criteria 

Fatigue-life prediction methods based on Lundberg-Palmgren theory are available to size 

traction drives. These methods are well established for determining life ratings of rolling-element 

bearings. However, they tend to underestimate the fatigue life obtained with today’s bearings. As a 

result, life-adjustment.factors have been identified (ref. 55) to more accurately reflect the longer lives 

attendant with improved bearing materials, design, and manufacturing techniques. It is anticipated 

that many of these same life factors are also applicable to traction drives, but endurance test data are 

needed to corroborate this and to identify what life adjustment factors, if any, are needed. 

The upper limit of power transfer for traction drives is relatively unexplored. Drives having 

power capacities of thousands of kilowatts can theoretically be constructed. Not enough is known 

about the practical thermal limits of traction contacts or their ability to tolerate transient slips under 

unusual circumstances. The fact that gear teeth regularly experience momentary slips of 30 or 40 

percent on entering or leaving mesh suggest that transient slip values can be remarkably high. 

Another important task is the establishment of a consistent, universally accepted design 

standard for traction drives, similar to those that currently exist for gear drives, bearings, and other 

mechanical components. The design guides of Wernitz (ref. 27) and Kraus (ref. 97), published about 

20 years ago, were most welcome contributions to the field, but our basic understanding of traction 

phenomena has greatly improved during the interim. A clear, practical, and comprehensive guide for 

the design and assessment of traction transmissions would benefit the designer and user alike. The 

bulk of such a guide can probably be assembled from the wide spectrum of information that 

currently exists. 

Manufacturing 

In principle, traction drives should be relatively inexpensive to make in large quantities. 

Manufacturing techniques are essentially the same as those used in the production of rolling-element 

bearings. To date, traction drives have been produced in relatively limited quantities or on special 

order and as a result have not yet realized their low cost potential. 

143 
Summary 

The evolution of tractiondrive technology has been traced over the past 100 years. Some of the 

more prominent events in the development of traction drives appears in table 11. This list is by no 

means intended to be comprehensive but, rather, to give the reader some appreciation of the scope of 

activities leading to traction drives of today. 

The earliest of traction drives, constructed of wood, leather, or fibercovered disks running in 

dry contact, found first use in factory equipment before the turn of the century. The ability of 

traction drives to smoothly and efficiently vary speed made them natural choices as main 

transmissions for several early vintage cars, such as those produced by Cartercar, Sears, Lambert, 

and Metz. Apparently, durability problems with the soft-materialcovered disks used in these drives 

foreshortened their commercial success. In the 1920’s traction drives equipped with oil-lubricated, 

hardened, steel rollers started to appear. These drives had much greater power capacity, and by the 

1930’s several industrial adjustable speed traction drives were being marketed both here and in 

Europe. About this time, there were several projects to develop toroidal traction car transmissions, 

notably the Hayes and later, the Perbury efforts in England and the General Motors’ work in the 

United States. In the 1940’s modem lubrication and fatigue theories for rollingcontact elements were 

developed, and these were later adapted to the design of traction drives. In the 1950’s work began on 

identifying fluids with high traction properties and experiments on how these fluids actually behaved 

within the traction contact. A basic understanding was also obtained on how the fluid film within the 

contact was compressed into a thin, stiff, tenacious solid-like film across which considerable torque 

transfer could safely occur. By the end of the 1960’s highquality bearing steels and traction fluids 

were commercially available. The power capacity of traction drives using the new steels and fluids 

virtually tripled. In the 1970’s improved traction models and fatigue-life prediction methods were 

developed. This all lead to the development of a new generation of traction drives-drives with a 

bright potential role to play in the power transmission industry. 

In 1980 the Power Transmission and Gearing Committee of ASME took a major step in 

recognizing the potential viability of traction drives by establishing a subcommittee to follow the 

developments in the technology for these transmissions. A primary function of this subcommittee is 

to assist in the dissemination of technology related to traction drives and to foster their potential use 

in industry. 

TABLE 11. - LIMITED CHRONOLOGY OF TRACTION DRIVE DEVELOPMENTS 

Date 

1870 

1877 

1906 

1921 

1923 

1926 

1928 - 1933 

1935 

1939 

1947 

1949 

1951 

1955 - 1962 

1955 

1957 

1965 

1968 

1976 

1978 

1980 

- 

1962 - 1968 

- Event 

Wooden friction drives used on wood-working machinery Hunt toroidal friction drive patented Friction drive equipped Cartercar introduced 

Automatic contact loadin mechanism patented by Erban in Germany Oil lubricated, Arter iniustrial drive marketed in Switzerland 

Carter analyses creep between locomotive wheel and rail 

General Motors Research Labs. road tests toric traction 

transmission 

Austin Motor Company offers Hayes Self-Selector toroidal 

transmission on Austin SixteenEHD lubrication theory advanced by Ertel and, later, Grubin 

Lundberg and Palmgren develop fatigue theory for roll ing elements 

Kopp Ball Variator comnercially introduced in Switzerland 

Clark, et al., theorize oil solidifies in EHD contact 

Braunschweig University experiments on traction contact phenomena 

in German 

Vacuum-me ted bearing steels introduced 

Lanels experiments to identify high traction oils in England Hewko’s investigations into traction contact performance Nasvytis devises fixed-ratio, multiroller planetary drive Monsanto and Sun Oil introduce comnercial traction fluids 

Coy, et al., apply Lundberg-Palmgren fatigue life theory to traction drive contacts 

Johnson and Tevaarwerk traction model applied to the design of 

traction drives 

ASM establishes traction drive comnittee 

T 

144 
. 

I 

Field experience has been gathered by industrial tractiondrive manufacturers, some of whom 

have been making traction drives for more than 40 years. However, tractiondrive technology is 

relatively young. The latest generation of traction drives has reached a high level of technical 

readiness. As these drives find their way into industrial service and as work continues in the 

laboratories, further improvements and increased usage of these drives can be expected. 

Concluding Remarks 

In reviewing tractiondrive developments through the years, it becomes evident that relatively 

few developments resulted from basic research. Technical advancements were largely due to 

individual organizations striving to perfect their own specific systems. As a consequence, meager 

fundamental information was left behind for those who followed. Of course, as this review has 

attempted to point out, there have been notable exceptions to this, but, by and large, relatively little is 

known about the reasons for the many commercial diasappointments that occurred along the way. 

It is encouraging, at least to this writer, that there currently seems to be greater communication 

among people working in the field. Admittedly, their numbers are miniscule compared with those 

.working in the bearing and gear disciplines. Nonetheless, tractiondrive technology of late seems to 

be making good progress. This is due, in no small measure, to much of the fine work that has been 

performed on tribological fundamentals, common to all rollingelement components. It is likely, in 

the years to come, that traction drives will find an increasingly larger role to play in the power 

transmission industry. 

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145 
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56. Coy, J. J.; Loewenthal, S. H.; and Zaretsky, E. V.: Fatigue Life Analysis for Traction Drives with Application to a 

57. Rohn, D. A.; Loewenthal, S. H.; and Coy, J. J.: Simplified Fatigue Life Analysis for Traction Drive Contacts. J. Mech. 

58. MacPherson. P. B.: The Pitting Performance of Hardened Steels. ASME Paper No. 77-DET-39, Sept. 1977. 

59. Soda, N.; and Yamamoto. T.: Effect of Tangential Traction and Roughness on Crack Initiation/Propagation During 

60. Tallian, T. E.; Chiu. Y. P.; and Van Amerongen, E.: Predictions of Traction and Microgeometry Effects on Rolling 

61. Bamberger, E. N.: Materials for Rolling Element Bearings. Bearing Design-Historical Aspects, Present Technology, and 

710837, Oct. 1971. 

NASA TN D-8124, 1976. 

no. 3, July 1979, pp. 251-257. 

Mech. Eng., 1971. 

Toroidal Type Geometry. NASA TN D-8362, 1976. 

Des., vol. 103. no. 2, Apr. 1981. pp. 430-439. 

Rolling Contact. NASA TM-81608. 1981. 

Contact Fatigue Life. J. Lubr. Tech., vol. 100, no. 2, Apr. 1978, pp. 156-166. 

Future Problems, Am. SOC. Mech. Eng., 1980, pp. 1-46. 

146 

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69. Magi, M.: On Efficiencies of Mechanical Coplanar Shaft Power Transmissions. Chalmers University, Gothenburg, 

70. Hewko, L. 0.: Roller Traction Drive Unit for Extremely Quiet Power Transmission. J. Hydronautics. vol. 2, no. 3, July 

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75. Kannel. J. W.; and Walowit, J. A.: Simplified Analysis for Tractions Between Rolling-Sliding Elastohydrodynamic 

76. Poon, S. Y.: Some Calculations to Assess the Effect of Spin on'the Tractive Capacity of Rolling Contact Drives. Proc. 

77. Lingard. S.: Tractions at the Spinning Point Contacts of a Variable Ratio Friction Drive. Tribol. Int., vol. 7, Oct. 1974, 

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147 E-817-8 
Namnill Aeronauucs and -e Adminislrabon 

1. Report No. 

NASA TM-101430 

Report Documentation Page 

2. Government Accession No. 

9. Security Classif. (of this report) 

Unclassified 

7. Author@) 

Donald H. Buckley, William R. Jones, Jr., Harold E. Sliney, 

Erwin V. Zaretsky, Dennis P. Townsend, and Stuart H. Loewenthal 

20. Security Classif. (of this page) 21. No of pages 22. Price., 

Unclassified 150 A07 

9. Performing Organization Name and Address 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135-3191 

12. Sponsoring Agency Name and Addm 

National Aeronautics and Space Administration 

Washington, D.C. 20546-0001 

3. Recipient's Catalog No. 

5. Report Date 

6. Performing Organization Code 

8. Performing Organization Report No. 

E4535 

10. Work Unit No. 

505 

11. Contract or Grant No. 

13. Type of Report and Period Covered 

Technical Memorandum 

14. Sponsoring Agency Code 

15. Supplementary NotPrepared for Seminar F-107 at the 1985 International Trade Fair, Cleveland, Ohio, October 18, 1985. This 

report is a compilation of papers presented at other meetings and published elsewhere. 

IS. Abstract 

This is a compilation of papers on the subjects of lubricants and lubricated components. The first paper addresses 

the fundamentals of tribology (the study of adhesion, friction, and wear). This is followed by a survey of the 

technology and science of boundary lubrication with liquid lubricants; another paper describes materials and 

methods used in solid lubrication practice. Conventional solid lubricants such as graphite and molybdenum 

disulfide are described, followed by a discussion of novel solid lubricants for very high temperature applications. 

In addition, there are three papers on the subject of lubricated mechanical components. The scope of these papers 

includes the lubrication and design of high speed rolling element bearings, high speed gears, and traction drives. 

17. Key Words (Suggested by Author@)) 

Tribology fundamentals; Boundary lubrication; 

Solid lubrication; Rolling element bearings; 

Gears: Traction drives 

18. Distribution Statement 

Unclassified - Unlimited 

Subject Category 39 

'For sale by the National Technical Information Service, Springfield, Virginia 22161 NASA FORM 1626 OCT W