Friction, Lubrication, and Wear Technology (1997) Part 10 potx

The variation of brake lining specific wear rate with brake drum or disk cast iron temperature is shown in Fig.. Material exchange between the cast iron transfer layer and the semimet li

Brake Rubbing Speed Effects. Organic brake linings show little variation in their specific wear rate with braking speed, when tested from low initial cast iron temperatures There typically is a slight rise of wear rate below a rubbing speed of 2 m/s (6.5 ft/s) Brake asperity "flash" temperatures are known to vary primarily with speed, as has been described by Blok (Ref 2) and, more recently, by Lim and Ashby (Ref 3) Above 1 m/s (3.3 ft/s), asperity temperatures appear to range from 1000 to 1100 °C (1830 to 2010 °F) It is presumed that the variation of lining wear rate and the associated variation of friction level are related to this flash temperature transition At very high rubbing speeds, the lining wear rate increases This increase is greater when the initial brake drum or disk temperature is high It is presumed that this speed effect is simply the result of higher interfacial temperatures Model wear data, presented later, support this presumption

Semimet friction materials also exhibit unique behavior with speed At rubbing speeds below 2 m/s (6.5 ft/s), semimet lining wear rates also increase, but to a significantly greater extent than do the organic linings Thereafter, the semimet brake linings provide a nearly constant specific wear rate with rubbing speed, until a transition condition is reached Higher speeds then generate much higher lining wear rates (to 100 times, or more) During full brake dynamometer testing of semimet brake linings, it was found that four brake stops from 160 km/h (100 mph) produced as much lining wear as over 500 brake stops from 50, 65, 100, and 130 km/h (30, 40, 60, and 80 mph)

There are differences among commercial semimet brake lining formulations in terms of rubbing speed transition values Higher transition speeds were found with a semimet lining that contained a small amount of para-aramid pulp It is conjectured that this high-strength thermoplastic material provided enhanced near-surface brake lining strength and helped to prevent "friction welding" wear of the lining to the brake rotor

Brake Temperature Effects. The variation of brake lining specific wear rate with brake drum or disk cast iron temperature is shown in Fig 1 for four representative lining classes The friction materials are divided into semimet (SM), light-duty (LD) organic, heavy-duty (HD) organic, and original equipment (OE) organic classes Because asbestos and NAO materials have overlapping wear properties, they are not separated

Fig 1 Temperature effects on brake lining specific wear rates

Semimet linings also exhibit a typical behavior that requires some explanation These friction materials establish a transfer layer to the cast iron during break-in and initial service Until established, the semimet wear rates are several times the normal rate

This cast iron conditioning process occurs faster with higher surface temperatures, higher rubbing speeds, and higher unit pressures It also varies somewhat among the different semimet lining formulations, and even with differences of cast iron countersurface, such as texture, residual stress, and oxidation

Once the cast iron surface is conditioned, it is believed that a back-and-forth transfer of material takes place Material exchange between the cast iron transfer layer and the semimet lining surface helps to keep the measured lining wear rate low No direct cast iron contact or wear occurs, except for incidental scoring that is due to particulate contamination, after the transfer layer is established This is further described in the section on brake drum and disk wear

Brake Usage Severity Effects. Figure 2 shows the wear life behavior for several classes of brake linings under different severities of usage It can be seen that no friction material type is best for all usage conditions Inexpensive aftermarket (AM) friction materials have acceptable wear lives under light usage conditions, but wear rapidly under more stringent conditions Heavy-duty (HD) friction materials have a lower wear life variation with usage severity than most automotive materials, but are superior in wear behavior only for severe-duty usage Sintered metallic friction materials also may have this wear characteristic, with low variations of wear rate at different usage conditions Semimet (SM) brake linings are best in the middle range, but poor in very light duty conditions Original equipment (OE) brake linings have a broad range of acceptable wear life From Fig 2, it should be clear that friction material wear life values are essentially meaningless unless usage conditions are specified It is not possible to select the optimum friction material for wear without knowledge of the customer brake usage distribution

Fig 2 Brake lining wear life versus usage severity

Brake lining Wear Modeling

Archard's equation is applicable to automotive brake systems only at very low rubbing speeds and component temperatures Consequently, this linear relation between volume of wear and the product of sliding distance and applied load is not generally applicable for brake wear modeling Forcing friction material wear data to fit polynomials of speed, rubbing distance, and load terms has been tried, but with very limited applicability

Rhee (Ref 4) and others have related measurements on friction material wear with the brake cast iron temperature These data indicated an essentially constant wear rate for low temperatures and a nearly exponential one at elevated temperatures

Lining Cure Effect. Weintraub and Bernard (Ref 5) derived a model equation which showed that some pyrolytic gas chromatography (PGC) peak areas of a brake lining resin behaved in an orderly manner with curing time and temperatures Simple laboratory-prepared formulations were used They also showed that laboratory friction material wear rates for these simple brake linings were linearly related to these PGC peak areas This work led to a simple, second-order Arrhenius-type model that was expressed as:

where V is the specific wear rate (wear volume per unit frictional work), A is the low-temperature lining wear constant, B

is a friction material wear constant, C is another friction material constant, and T is the average interface temperature

during braking

This equation has been quite effective in relating brake lining wear data with testing conditions for a broad range of materials, over most of the brake operating temperature range Model formulations and production friction materials had full-brake wear behavior that was found to be well-characterized by this relationship The production linings had several

organic components, making the applicability of this equation surprising Experimental values found for the constant C

appear to have functional significance For most friction materials, this activation-energy-related term has been determined experimentally to vary from 65 to 85 kJ/mol (16 to 20 kcal/mol) Such values might well be expected from bimolecular reactions in the wear process

Because the constant term A dominates at low operating temperatures, it is easily determined experimentally However, following some severe usage conditions, especially after high friction material soak temperatures, the value of A was

found to increase and subsequently remain at this increased value The increase appeared to be associated primarily with the friction material soak temperature, but also with some dependence on the time at temperature It does not appear to be simply related to the mean interface temperature during braking Thus, a better model expression would have the constant term replaced by one that includes prior thermal history This new term also could include particulate contamination factors for an even more sophisticated model

As the temperature increases, the thermally sensitive Arrhenius-type term increases in influence The temperature required in the above wear equation is the mean rubbing interfacial temperature This is not easily measured, but it can be calculated The readily calculable interfacial temperature rise for the brake application is added to the measurable bulk

cast iron temperature of the brake drum or disk Values for B and C can then be determined from experimental wear data

for temperatures, usually in the 150 to 350 °C (300 to 660 °F) range

Whether this wear equation is technically correct or not is not at issue It generally works well, and unique wear mechanisms appear to be involved when it does not To a pragmatic experimentalist, this has been quite acceptable

Brake Drum and Disk Wear

Gray cast iron is the dominant material for both brake drums and disk brake rotors Brake cast iron typically has a type A graphite, with a pearlitic matrix and low ferrite and carbide content Normal brake iron provides adequate damping and good resistance to thermal fracture in hard service

The wear of either the cast iron or other countersurface is not readily determined by the Arrhenius-type model relationship It appears that several mechanisms affect the specific wear rate of the cast iron These include abrasive, adhesive, and oxidative terms The abrasive terms include components from the abrasive content of the friction material, external contaminants, and abrasive particles that are "manufactured" at the rubbing interface The latter can result, for example, from vitrification (firing) of clays at the hot rubbing surface

Brake Lining Chemistry Effects. Some friction materials have constituents that are chemically active at the rubbing interface Such brake linings have been found to be extremely responsive to changes in the cast iron chemical composition Hatch (Ref 8) described the effect of titanium and rare earth oxide content on lining friction Examples of large brake effectiveness differences with the same brake lining, but different cast irons, were shown Regrettably, this article did not make clear that most friction materials do not provide any detectable difference in effectiveness, only the few that chemically interact with the cast iron

With one such disk brake lining, the in-service cast iron wear life was increased by a factor of 50 when the titanium content was increased from 0.02 to 0.04% The presence of small, hard particles in the cast iron from the titanium was credited for the improved wear resistance on this countersurface Titanium content control can be particularly effective in disk brakes, if parasitic drag of the brake linings causes local wearing of the rotor Cast iron machinability considerations limit the useful titanium content of 0.05%

Graphite Morphology Effects. The graphite size and shape were found to affect the cast iron and lining wear rates for many brake linings "Damped" gray cast iron, with its very large graphite flakes, has been used in some brake applications to decrease brake squeal However, this iron is weaker, and tends to have poorer wear resistance than conventional brake iron It also increases brake lining wear rates for some friction materials This presumably results from the cutting action of sharp iron edges found around the large graphite flakes Linings with hard resins and rigid matrixes were found to have lower wear rates with decreasing graphite size

Special cast iron brake test parts were made with very fine graphite structures and with about 0.04% titanium Using a hard and abrasive nonasbestos truck brake block, this permanent mold cast iron provided a 30% reduction in the brake lining wear rate, and had a 240% improvement in cast iron life, compared with a conventional brake cast iron

For the NAO heavy truck brake blocks in particular, the cast iron chemistry and graphite morphology can exert a strong influence on the countersurface wear life It appears that many nonasbestos materials cause the cast iron to become a more active member of the friction and wear couple Consequently, closer control of the cast iron may be required Each new friction material should be tested for cast iron sensitivity, to ensure acceptable service life

Normal Cast Iron Wear. Usually, a worn cast iron rubbing surface develops a satin gray appearance abrasiveness passenger car brake linings provide cast iron specific wear rates from about 0.2 to 1 mg/MJ (1.6 to 8 × 10-

The cast iron wear rate is sensitive to a number of factors An abrasive will wear the cast iron if it has higher hardness, higher melting-softening temperature, and sufficient particle size Litharge and barite are both softer than cast iron at room temperature With normal brake usage, barite will increase the cast iron wear, whereas litharge will not, because barite has a higher softening temperature than cast iron Even materials that meet the hardness and softening criteria may not produce severe cast iron wear, unless they are of a size that can cause abrasion

A brake lining matrix hardness varies considerably with formulation and with temperature for a given formulation Some brake linings soften significantly, and can produce cast iron wear rates as shown in Fig 3, trace A Other linings have a nearly constant matrix hardness, and result in wear like that of trace B The cast iron wear rate increases at elevated temperatures, mirroring the lining wear rate When abrasives are evenly distributed throughout the brake lining, cast iron

wear rates directly follow the lining wear rates If external abrasives are present and dominant, wear behavior like that of trace C results The decrease of cast iron wear at higher usage temperatures then appears to result from softening of the brake lining matrix and abrasive clearance from higher lining wear rates

Fig 3 Cast iron specific wear rates versus brake temperature See text for discussion

Local Cast Iron Wear. Brake rotors can generate local brake lining contact zones when there is a significant axial runout Figure 4 shows the axial runout and thickness variation (TV) of a rotor that had minimal brake usage, but developed a thickness variation from parasitic wear (unintended brake dragging) during highway driving conditions Road dust contamination aggravated the wear Similar wear patterns have been seen with semimet brake linings, because of their magnesium oxide abrasive content

Fig 4 Rotor axial runout and thickness variation from local drag wear

Highly localized cast iron wear is found on the inboard face of the rotor, centered at a site slightly in advance of the maximum runout position The resultant brake pedal vibrations from this wear were sufficient to require the remachining

of the rotors after only 17,000 km (10,560 miles) of dusty highway service Often, a small increase of cast iron matrix hardness and hard particle content will reduce the local cast iron wear rate greatly

Extreme scoring of cast iron can result from an inadvertent contamination of the lining by abrasives For example, silicon carbide was a "tramp" impurity in the synthetic graphite used in a semimet brake lining formulation After 2000 km (1240 miles) of customer service, one rotor face was deeply scored over the entire rubbed surface The opposite rotor face showed no scoring, with the original grinder marks still evident on the surface Both brake linings were of the same semimet formulation, but from different production batches, one with silicon carbide and the other without Laboratory testing showed the contamination level at a few tenths of a percent

Hot spotting of the cast iron can produce local wear at the heated sites If martensite is formed on the cast iron surface, local high spots will result Martensite and hot spotting have been addressed by Anderson and Knapp (Ref 9)

External Abrasive Effects. Dust or splash shields can be used to reduce disk lining and rotor cast iron wear from particulate contamination, but attendant restricted brake cooling may increase lining wear rates External contaminants can be kept from drum brakes by labyrinth seals or shrouding with a small loss of cooling efficiency However, retention

of normal wear products and accumulated rust particles within the brake sometimes then contribute to increased wear rates With riveted linings, the cast iron wear may be heavily biased along the path of the rivet holes, particularly those at the leading edge of the linings Careful testing and intelligent compromises may be needed to balance cooling and contamination effects

Transfer coatings onto the cast iron can be generated by semimet brake linings, as described in the section on brake lining wear These coatings can vary from a few tenths of a micron to 40 m (1.6 mils) or more Once formed, the wear

of the original cast iron becomes zero, except from local scoring that is due to abrasive action However, the transfer coating is by no means static

With extended operation at low temperatures and especially for rubbing speeds below 2 m/s (6.5 ft/s), the coating will deplete This appears to result from a preferential transfer to the friction material, because very low lining wear rates are measured during the period when the coating is depleting Once depleted, the brake lining wear rate increases by a factor

of about 4 The cast iron then wears at a substantial rate, and its appearance changes from a dark blue-black coloration, with a matte finish, to a shiny gray With time, the cast iron usually develops uneven wear across the rubbing path During near-transition usage, the cast iron coloration includes bands of brown, blue, and purple

With continued low-temperature, low-speed usage, the cast iron wear mass removal rate has been found to approach that

of the friction material If these conditions continue, the disk brake linings will wear out at about 20% of the expected mileage, and the rotors will have become unserviceable from the high and generally uneven wear Normally, such high wear rates are associated with severe usage In this case, a change of wear mechanism causes high wear under very light-duty, low-temperature usage

If the cast iron wears under light-duty conditions, but not under hard service, one might redesign the brake to ensure higher temperatures This measure is acceptable, within limits Continued high temperature and hard brake usage cause another unique condition The transfer layer may become so thick that it locally delaminates from the cast iron rotor This occurs when the transfer layer thickness is about 30 m (1.2 mils) Figure 5 shows a cracked and incipient failure state for the semimet transfer layer on a medium truck disk brake rotor that had seen severe service on a mountainous test route Local high spots, around 1 mm (0.04 in.) in diameter, signal these delamination sites The light bands near the center of the photograph are regions where the coating has broken away locally

Fig 5 Semimet transfer layer cracks and blisters

Figure 6 shows a more severe case, with the same semimet lining and brake usage The cast iron is visible at the bottom

of the recesses, where the transfer coating has flaked off Machining marks could still be seen at these sites Based on a simple scratch test, the coating was found to be quite adherent Presumably, the delamination process resulted from high thermal strains at the coating-cast iron interface

Fig 6 Semimet transfer layer delamination

Design analysis of brakes for semimet linings thus appears to involve assurance that light-duty brake usage will not produce rotor wear and severe usage will not result in transfer layer delamination Consequently, optimal design would prevent these usages in most operating situations, and would result in rare occurrences of either wear type The greater the variation of usage severity in customer service, the greater the possibility of such cast iron wear issues

Toxicity of Brake Wear Debris

When brake wear debris is considered, asbestos fiber is a common concern However, several studies of brake lining wear emissions have shown only about 0.03% asbestos in the airborne wear debris The low emission results from the high frictional contact asperity temperatures Chrysotile asbestos converts to forsterite at about 800 °C (1470 °F), well below brake flash temperatures Organic brake lining fiber emissions averaged 0.7 m (28 in.) in length Therefore, the vast majority of the airborne asbestos did not meet the definition of a fiber by the Occupational Safety and Health Administration, the World Health Organization, and other authorities Epidemiological data for automotive brake mechanics have shown no detectable excess cancer risk The Environmental Protection Agency, using computer models for fiber exposure and cancer risk, predicted about 15 excess cancer cases a year from all U.S brake wear

Semimet brake wear emissions appear to be of minimal toxicity Fibers from the wear of other nonasbestos brake linings have not yet been collected and measured These include mineral wool, glass fiber, slag wool, titanate fiber, phosphate fiber, carbon, and para-aramid fiber

Although lead was removed from most brake linings by 1980, barium, usually in the form of barium sulfate (barite) powder, is often found Such heavy metal wear particles may affect chest x-rays of brake mechanics Brake usage generates some fused-ring aromatic material, but in miniscule quantities, like those in lettuce Until complete and definitive brake wear toxicity studies are completed, it would be prudent to minimize dust exposure from any brake wear debris

Automotive Brake Frictional Characteristics

Brake linings inherently have some performance attributes, such as fade and fade recovery, but full evaluation of a friction material requires that it be installed into a full brake system Vehicle brakes are required to operate under a wide range of conditions, from hard braking with a heavily loaded vehicle on a steep downhill slope to minimal brake usage on interstate highways Vehicle brakes should be highly reliable and minimally affected by temperature, water, or other contaminants Brake lining friction must be consistent throughout the life of the material

Brake Design Basis. Brakes are designed primarily on the basis of wear, pedal travel, brake system stability, and effectiveness properties Brake effectiveness is defined as the ratio of the brake friction torque to the applied force This term is used, rather than friction coefficient, because brake geometry factors often make the brake torque not linearly related to friction level For example, effectiveness must be used to describe frictional performance of drum brakes, because of their large self-actuation characteristics

Self-Actuation. Frictional forces acting on the shoes of a simple drum brake may cause it to be further loaded against the drum, increasing its effectiveness Such shoes are referred to as leading Friction force on a trailing shoe causes it to oppose the application force, decreasing effectiveness The increase in brake shoe loading, which is due to friction and geometry effects, is referred to as self-actuation

Although the geometric considerations of self-actuation influence brake effectiveness, it also is influenced by frictional properties of the brake lining Figure 7 shows design effectiveness curves for different brakes as a function of friction coefficient Assuming a nominal friction coefficient of 0.4 for each 1% change in friction coefficient, the brake effectiveness changes by approximately 3.5% with a duo-servo drum brake (two leading shoes, coupled), and by 2.6% on

a leading-trailing drum brake, but the change is just 1% for a nonservo disk brake

Fig 7 Effectiveness versus friction coefficient for several brake types

Design Factors. In addition, the design of brakes can affect the degree to which friction material wear and thermomechanical properties influence brake effectiveness A large change of brake effectiveness is possible, for the same value of friction coefficient, simply because of a change of lining pressure distribution Figure 8 shows the change

of effectiveness on a leading shoe that results from different pressure distributions For a given value of friction coefficient, the effectiveness may vary by a factor of 4 as the pressure distribution varies from center-biased to end-biased On a drum brake, the lining pressure distribution is affected by the stiffnesses of drum, shoes, and brake linings, in addition to thermal distortions, wear patterns, and prior usage history effects

Fig 8 Effectiveness variation with pressure distribution, based on type of leading shoe

The brake designer can achieve the same brake effectiveness by several combinations of lining geometry, brake shoe geometry, and lining coefficient of friction Thus, changing friction material type (as from asbestos-organic to one of the nonasbestos types) may provide significant effects on brake system performance, because of different coefficients of friction, different thermal expansion coefficients, and different compression stiffnesses

Many brake design parameters are interrelated For example, a large value for lining thermal expansion may require added running clearances to avoid parasitic drag This results in greater brake actuator travel, which may then necessitate further brake system redesign A complete introduction to brake design is beyond the scope of this article The preceding discussion was presented to introduce design issues and to illustrate the virtual impossibility of simple substitutions

Brake Frictional Performance

Some friction material effectiveness measures that are evaluated during new vehicle brake system qualification experiments include: burnish and green performance, fade resistance, fade recovery, delayed fade, speed sensitivity, effectiveness drift, and environmental sensitivity Each of these important performance characteristics is further described below These are the most important of a much larger set of brake performance attributes that are evaluated during brake system development

Green Effectiveness. When a friction material is new, it is unburnished and its contact with the drum or disk depends

on many manufacturing tolerance stack-ups With a new brake assembly, effectiveness often is lower and more variable than after the friction material is fully burnished This initial behavior is called the green effectiveness

The initial brake lining friction values may vary because of surface texture effects, surface contamination effects, contact pressures, and brake lining composition Particularly with drum brakes, the green effectiveness also depends on the initial contact geometry Design dimensions are chosen to provide lower initial effectiveness from the contact geometry This minimizes the possibility of the overeffectiveness of new brakes, which is due to either rust or some other contaminant that increases initial friction

Burnished Effectiveness. After the drum or disk has contact over the full rubbing surface, and when a stable char layer has formed over the entire brake lining contact area, the brake is said to have been fully burnished Typically, the brake performance has stabilized by the time that 85 or 90% of the brake lining has contacted the brake drum or disk An exception to this exists for the semimet brake lining Semimet frictional behavior is discussed separately, because it has several unique characteristics

Fade Resistance. Brake fade refers to a loss of brake effectiveness, generally as the result of excessive brake temperatures However, five types of brake fade can occur:

• Thermal, which is due to high brake bulk temperatures

• Delayed, which is due to resin migration during cooling

• Blister, which is due to near-surface lining blisters

• Flash, which is due to high-speed, high-load braking

• Contamination, which is due to partial lubrication from water or oil

Poor brake fade behavior could be exhibited by high-quality materials, if used is inappropriate brake applications Brake lining fade behavior is meaningful primarily in the context of a particular brake lining, brake, and vehicle usage situation Depending on the brake usage sequence and lining fade response behavior, a driver may either sense impending fade as a gradual increase of brake pedal force requirement or he may not have any warning Good brake lining choices provide a gradual and progressive loss of effectiveness to warn of impending brake fade Good applications do not provide the driver with surprises

Fade recovery refers to the ability of the friction material to quickly regain normal brake effectiveness after thermal fade As the brake disk or drum cools, the friction level should return quickly to the prefade level

Delayed fade is a phenomenon that can occur with some drum brake friction materials During the fade recovery, brake effectiveness may drop unexpectedly, causing a temporary, but pronounced, increase of brake pedal force requirement A delayed fade is insidious in that it is unexpected It occurs well after a period of hand brake usage and usually with no warning signs

Blister Fade. New brake linings may contain volatile material from fabrication, which, if not released by the end of the burnish process, could cause high internal gas pressures upon rapid heating, as during a hard brake application In some situations, a near-surface blistering results in a rapid, brief loss of brake effectiveness Friction is lost because of excessive contact pressures at the blister sites and from the evolved gases A characteristic of blister fade is brevity Effectiveness is lost for a few seconds during a hard brake application, and then returns to near-normal Repetition of the hard brake application will not produce a second blister fade, because the volatile material has been eliminated from the near-surface region

Flash fade is related to the blister and green fade, but occurs only at very high brake power levels, usually at very high speeds The rapid decomposition of near-surface organic constituents produces a gas-pressure-lubricated braking surface The brake lining friction may not be low, but the evolved gas pressurizes the friction material to counteract the applied force High surface area brakes, and those with high organic contents, are most vulnerable Prior brake usage at moderate-to-high brake lining soak temperatures reduces flash fade severity

Contamination Fade. Water, oil, or a combination of these on the surface of the brake lining or brake drum/disk can generate an elastohydrodynamic fluid film that effectively makes a bearing from a brake Different friction materials have

different porosity, compliance, and wear characteristics, and thus may be quite different in sensitivity to contamination High surface area brakes, such as drum brakes, are more prone to contamination fade However, even automotive disk brakes can exhibit such a fade, if saturated by either water or oil/water This effect is similar to that from oil/water films

on the road surface, affecting tire friction

Speed Sensitivity. All brake linings exhibit different frictional behavior as the rubbing speed changes The speed sensitivity varies with material type and composition It also varies with temperature, pressure, and prior usage history To ensure proper brake balance over a wide range of stopping speeds, the brakes on each axle should exhibit similar effectiveness characteristics

In general, brake effectiveness decreases with increasing speed Disk brakes provide less speed spread, or difference in brake effectiveness at different braking speeds, than do drum brakes This is primarily due to the low servo factor of disk brakes Duo-servo drum brakes, with high self-energization, are the most sensitive to speed Good front-to-rear brake balance at all speeds is difficult to attain on vehicles with front disk brakes and duo-servo rear drum brakes Highly speed sensitive disk brake lining need to be paired with low-speed-spread drum brake linings to compensate fully for the 3- to 4-fold difference in brake servo factors

Environmental Sensitivity. Contaminants such as water, oil, dust, and rust alter brake effectiveness For example, contaminant oil films of 0.2 m (8 in.) thickness can raise the brake friction coefficient by 55% at low vehicle speeds This can make some drum brakes virtually self-locking However, at 0.5 m (20 in.), an oil film can lower friction by elastohydrodynamic action Because of low oil tolerance and high sensitivity, even an oily fingerprint can cause significant effectiveness change

Wet Friction. The performance of vehicle brakes when wet also represents a safety concern Disk brakes usually are less affected by water than are drum brakes, largely because of their lower inherent servo factor However, both disk and drum brakes may show large effectiveness losses when wet A small amount of oil, together with water (as with tires) increases the lubricating action

When wet, some friction materials provide a greater loss of brake effectiveness than others These usually also take a considerably longer time to recover friction capability after wetting Permeability, homogeneity, and compression stiffness are some of the brake lining properties that help determine wet friction response A complete understanding of this behavior is not known Full brake dynamometer and vehicle tests are used to establish the wet friction behavior of brake systems

Moisture Sensitivity. Friction materials are typically somewhat porous, fiber-reinforced composites that are capable of absorbing atmospheric moisture when a vehicle is parked, such as overnight This moisture has been shown to change break effectiveness temporarily for some friction materials Glass and para-aramid ten to increase friction when moist Other materials tend to lower friction in moist environments

Drum brake linings also change in thickness with time when exposed to high-humidity conditions When the brake lining ends thicken, relative to the center, brake effectiveness increases

Rust Effects. Another moisture effect results from rusting to the cast iron disk/drum brake lining surface when a vehicle

is parked for some time, causing "stiction." Under adverse environmental conditions, it is possible for some friction material to rust-bond to the cast iron A substantial torque can be required to break this bond Thereafter, the rusted surface of the brake drum or disk may generate an uneven brake torque with brake angular position

Rusting of the brake drum or rotor can occur independent from rust-bonding In this situation, the brake effectiveness tends to be high at low speeds and low-to-normal at higher vehicle speeds Whether rust prevails or not depends on the relative abrasiveness of the brake lining and rotor surface, as well as the brake usage conditions

Although some of the above characteristics can be investigated in the laboratory, environmental testing is required to establish acceptability of friction material formulations in full vehicle service usage The ideal brake assembly exhibits little or no moisture sensitivity When present, it usually persists for a few brake applications, and then disappears when brake heat drives the moisture from the brake lining or when wear removes the surface rust A small abrasive content in a brake lining may be beneficial in removing minor rusting from the brake drum or disk Brake lining formulations also can contain agents to control minor rusting in service

Effectiveness Drift. To ensure consistent vehicle braking performance, the brake effectiveness characteristics should

be stable throughout the life of the brake linings The brake effectiveness of some friction materials can deteriorate with accumulated usage history, time, and wear The formation of a surface glaze on some brake linings, during prolonged light-duty usage, has been associated with a loss of effectiveness Linings with low wear rates tend to glaze more readily that do those with poor wear life

Semimet Frictional Behavior

Resin-bonded metallic, semimet brake linings have sufficiently different performance characteristics that it is simpler to describe them separately Disk brake effectiveness with semimet brake linings is virtually constant with normal usage pressures and temperatures, for speeds ranging from 50 to over 200 km/h (30 to over 125 mph) Effectiveness drops slightly below 50 km/h (30 mph)

Thermal Fade. Semimet linings tend to be nearly fade-free When thermal fade is experienced, it occurs as a progressive loss of effectiveness with decreasing speed Significant brake fade primarily occurs at the end of the stop

Environmental Sensitivity. Semimet linings are prone to have a 50 to 60% loss of effectiveness after exposure to moist air for one hour or more This becomes the limiting condition and determines the amount of power brake boost required with most semimet brake linings (most other brake lining types are limited by initial fade behavior) It also restricts the available choices for compatible drum brake linings

Water Sensitivity. Water affects many friction materials greatly Semimets are minimally affected unless the water is accompanied by an oil film with a thickness of 1 m (40 in.) or more In that case, the frictional effect is similar to other brake linings However, disk brakes run warm enough to dry quickly, so water effects usually are neither critical nor long lasting

Laboratory and Vehicle Brake Evaluation

Laboratory specimen testing machines are commonly used to characterize and audit the quality, or sameness, of friction materials, using specimens from full brake linings The advantages of using laboratory systems for evaluation include automated testing, careful control of operating conditions, and more rapid measurement of brake lining characteristics In addition, laboratory evaluations that utilize a specimen test are less costly than full-scale brake dynamometer or vehicle experiments

Several different laboratory machines are used to determine brake lining properties Three commonly used laboratory test machines are described below

The friction assessment screening test (FAST) machine was developed for rapid "fingerprinting" of friction material specimens The FAST machine is used primarily for routine brake lining quality control testing Although this machine reportedly has been used to perform friction material screening tests, it never has been recommended as a substitute for full-scale brake evaluations

A friction materials test machine is also used for quality-control testing The Society of Automotive Engineers (SAE) has designated a recommended practice (SAE J661a), used in conjunction with this machine, that is classified as a quality-control test procedure It is used more for periodic quality-control surveillance testing, as opposed to routine production batch testing Based on test data from the SAE J661a procedure, a brake lining friction rating specification (SAE J866a) has been used by some states to classify and regulate friction materials However, this rating system has been shown to be clearly inadequate for meaningful comparative testing with different types or classes of friction materials SAE J866a (revised in 1984) includes a caution against such uses

A full brake inertia dynamometer simulates vehicle braking by mounting a complete brake assembly to a large rotating inertia and drive motor Most are single-ended, and test a single brake assembly at a time Brake dynamometers place the brake assemblies in closed ducts, both to expedite cooling and to control smoke and smell Because faster cooling rates hasten testing, most dynamometer tests have much greater air flow, and resultant cooling rates, than are found in on-road testing Consequently, brake dynamometers are used mostly for controlled wear tests, basic effectiveness tests, initial fade/recovery tests, and parking brake tests

Full brake dynamometers are available in a range of sizes, with inertial equivalents from motorcycles to railroad locomotives Because there are no brake dynamometer standards, many are unique Dynamometer-to-dynamometer differences in test results can be significant, even for carefully matched linings and brakes that are tested to the same inertia loading As a result, most brake test engineers prefer comparative brake lining test data, obtained from a single brake dynamometer, to data from different brake dynamometers

Full brake dynamometers can closely replicate most vehicle in-service braking conditions, except for air flow over the brake and environmental conditions, such as water and dust contamination They are therefore invaluable for brake diagnostic testing, and can be excellent for initial brake lining screening tests However, they are not sufficient to fully evaluate brakes

Correlation of Laboratory and Vehicle Test Results. Numerous studies have been done to determine the degree

of correlation between laboratory friction material test results and actual vehicle test results In general, the only good analog of a vehicle brake is the brake itself Consequently, there are no specimen or scale test devices that consistently yield test data that correlate with full vehicle data This does not mean that such laboratory tests are useless, just that they should not be used to predict field performance behavior They can, and have, been used to screen friction materials with performance flaws, but require expertise in data analysis

Even full brake dynamometers are difficult to correlate with on-road vehicle brake performance, unless careful instrumentation and test controls are used It should be possible to correlate full brake dynamometer test results with on-road data without a need for questionable correction factors Although this has been tried by many, and published by a few, it has not as yet been verified by anyone

The testing of brakes involves evaluations under the widest possible sets of conditions, in order to find flaws and behavior inconsistencies Friction material formulation and testing both are more of an art than a science If accurate models of brake performance were available, then testing would be much simpler However, each friction material formulation is proprietary, and can be unique in many respects to other materials Consequently, brake test engineers use a combination

of standard performance test procedures to ensure adequate capacity Customer service fleet tests over the entire United States then help to determine usage and environmental stability characteristics

Prudent brake development engineers start with friction materials that have been fully screened by laboratory tests, and then run full brake dynamometer tests to further screen the candidate brake linings and to initiate brake hardware adjustments Only then are vehicle performance tests justified When these are acceptable, the more difficult, costly, and time-consuming fleet and traffic tests are initiated

Federal Braking Requirements and Other Brake Tests. As indicated, the qualification of vehicle brake systems and friction materials can involve numerous experiments to determine fade resistance, moisture sensitivity, wet friction, and other performance parameters On new vehicles, brake performance must meet Federal Motor Vehicle Safety Standards (FMVSS) 105-75 for hydraulic brakes and 121 for air brakes Other performance standards are determined by the vehicle and brake manufacturers Replacement brake linings have no legislated performance standards to meet

SAE-Recommended Practices. The SAE has developed about 20 recommended practices for checking the performance of brake linings and systems These standards cover automobile, truck, and trailer brake system tests using both vehicles and dynamometers

Prior to the adoption of FMVSS 105 and 121, the SAE procedures were intended to provide guidelines to brake system evaluations The passage of FMVSS 105 as a requirement for brake system certification shifted the emphasis of these SAE procedures to the role of supplementary tests that could be used to further qualify vehicle brakes and braking systems The majority of the SAE brake codes have been directed toward automobiles and light trucks, although some (SAE J880 and SAE J9781) have been developed for heavy commercial vehicles

Friction material performance in vehicle brake systems can be qualified by meeting the requirements of the federal safety standards and being evaluated by the SAE recommended practices, along with brake component and vehicle manufacturer standards For new vehicles, compliance with federal and state motor vehicle safety standards is only a starting point for the acceptance of a brake system No federal standards exist today for replacement brake linings

References

1 S.K Rhee and P.A Thesier, "Effects of Surface Roughness of Brake Drums on Coefficient of Friction and Lining Wear," Preprint 720449, 1972222, Society of Automotive Engineers, 1937

2 H Blok, General Discussion, Proc Inst Mech Eng., Vol 2, 1937, p 222

3 S.C Lim and M.F Ashby, Wear-Mechanism Maps, Acta Metall., Vol 35, (No 1), 1987, p 1-24

4 S.K Rhee, "Influence of Rotor Metallurgy on the Wear of Friction Material in Automobile Brakes," Preprint

710247, Society of Automotive Engineers

5 M.H Weintraub and J.P Bernard, Chemical and Functional Responses to Brake Lining Cure Variations, SAE

Trans., Vol 77, paper 680416, 1968

6 M.H Weintraub, A.E Anderson, and R.L Gealer, Wear of Resin Asbestos Friction Materials, Adv in

Polymer Friction and Wear, Plenum Press, 1974

7 A.E Anderson, Wear of Brake Materials, ASME Wear Control Handbook, American Society of Mechanical

Engineers, 1980

8 D Hatch, Cast Iron Brake Discs, J Automot Eng., Oct 1972, p 39

9 A.E Anderson and R.A Knapp, "Hot Spotting in Automotive Friction Systems," presented at 1989

International Wear of Materials Conference and published in Wear, Vol 135, 1990, p 319-337

Friction and Wear of Tires

P.S Pillai, Goodyear Tire & Rubber Company

Introduction

TIRE WEAR is a complex phenomenon that occurs in the tire footprint area that is, the interface region between tire tread and pavement surface The forces required to support and control an automobile act in the footprint contact area Forces generated in the tire footprint allow the vehicle to change direction and speed of travel Production of these action and reaction forces generates frictional work in the footprint region The dissipation of this frictional work is the primary reason for tire wear Factors that affect tire wear are broadly classified in Table 1

Table 1 Parameters that affect tire wear

Tire-related factors Operating conditions Other factors

Tire type

Tread design

Tread composition

Tire construction features

Magnitude of inertial forces

Tire load Inflation pressure Speed

Driver vehicle interaction Vehicle maneuvering behavior

Pavement texture and condition Seasonal effects

Ambient temperature

Abrasion and Wear

Tire wear or tread wear refers to loss of tread rubber Abrasion of rubber refers to the removal of rubber particles from a testpiece (a small rubber wheel, for example) when rubbed against another hard object Thus, it would seem logical to assume that tread wear is related to rubber abrasion However, experiments have shown that the two processes do not exhibit a one-to-one correlation Informative review articles on rubber abrasion have been published (Ref 1, 2, 3, 4, 5)

Basic research on rubber abrasion and wear has been performed by Schallamach, Gough, Grosch, and other investigators (Ref 6, 7, 8, 9, 10, 11) Study of rubber abrasion is essentially limited to laboratory testing of model rubber specimens using abrasion machines; wear refers to tires in service Many authors have tried to use laboratory abrasion results to predict tread wear, but have had only limited success In the tire wear process, although te physical phenomenon is loss or abrasion of tread rubber, many factors tire type, design and construction parameters, material properties, vehicle maneuvering behavior, driver-vehicle interaction, weather, and so on influence wear rate Therefore, direct wear measurements are required to rank the wear rates of different tires The tread wear grade provided by tire manufacturers is

a comparative rating based on the wear rate of the tire when tested under controlled conditions on a specified government course Veith and co-workers have performed extensive experimental work on tire wear (Ref 12, 13, 14, 15) Veith has also published a comprehensive review of the topic (Ref 16)

Definition of Terms. Important terms related to tire wear are defined below

• Tread loss: Average cumulative loss in tread depth for all the grooves in the tire, expressed in

millimeters

• Rate of wear: Tread loss in millimeters per 10,000 km of travel

• Tread wear index: (Wear rate of reference or control tire)/(Wear rate of experimental tire) × 100

• Standard wear rate: mils per 1000 miles per 100 lbf cornering force at tire test temperature

• Camber angle: The vertical angle between the wheel plane and a line perpendicular to the pavement

(Fig 1)

• Slip angle: The horizontal angle between the tire circumferential midpoint plane and the direction of tire

motion (Fig 1)

• Cornering stiffness: Force per degree of slip angle

• Test severity: Function of tire force severity, pavement surface severity, or weather/temperature severity

• Tire force severity: Magnitude of cornering force

• Pavement surface severity: Micro- and macrotexture, blunt or harsh (qualitative ranking)

• Weather severity: Summer/winter temperature, rain, snow

• Severity index gradient: Rate of change of tread wear index with tire force severity

• Driving severity number (DSN): Sum of the squares of lateral acceleration per tire revolution (gi)

divided by the number of revolutions (N) normalized for vehicle weight:

DSN = (gi/100)2 (FzFz,R)2/N (Eq 1)

where Fz is the actual tire load and Fz,R is the rated tire load

Fig 1 Schematic showing tire footprint forces and tread distortion

Basic Mechanism. Figure 1 is a schematic of a tire making turn, the direction of motion, the slip angle, and the cornering and lateral forces are indicated The tire is skewed to the left by the slip angle The distortion of the tire

footprint generates a lateral force, mv2/R, so that the vehicle can make the turn (m is the mass on the tire, v is the instantaneous velocity, and R is the radius of curvature) The tread rubber is held to the road by rubber friction at the

leading edge of the footprint The forces acting in the footprint distort the tread area laterally The distortion increases in magnitude as the tread element travels from the leading edge toward the trailing edge of the lateral footprint When the local lateral distortion exceeds a certain limiting value, the frictional force cannot maintain the deflection and the tread element slides back toward the tire center plane The tread rubber stores elastic energy while in the distorted state This stored energy is partially converted into frictional work during the sliding domain The dissipation of this frictional work during the cornering mode is the source of free-rolling, untorqued tire wear

Tire Axis System. Figure 2 includes additional parameters (for example, tire forces, tire moments, and tire angles) that affect tire wear

Fig 2 Schematic showing angles, moments, and forces which comprise a tire axis system Source: Ref 17 Tire Wear Models

Theoretical Model. Using basic principles of physics, Schallamach and Turner (Ref 18) developed an equation for

even wear of tire treads in terms of average cornering force (F), cornering stiffness (C), tire resilience (R), and tread

rubber abradability ( ):

(Eq 2)

where A is the wear rate per unit distance Pulford (Ref 4) discusses some of the practical implications of Eq 2 Livingston

(Ref 19) extended the Schallamach-Turner model by partitioning the total energy stored into two parts a fraction stored

in the tire body as a whole and the remainder stored in the tread and obtained the following expression for tread life:

(Eq 3)

where h and h0 are final and initial tread height, respectively, x is distance traveled, and B and A are experimental

constants

Empirical Model. Experimental results have shown that the theoretical model expressed in Eq 2 must be modified as:

where k is a constant and the exponent n ranges from 2 to 4 The value of n depends on the slip angle, the nature of the

pavement, and carbon black particles (Ref 12) The Schallamach-Turner equation (Eq 2) indicates that force is the controlling factor in tire wear Gough (Ref 11) has also shown that cornering force is the major source of tire wear At equivalent force levels, cornering action causes higher energy loss in tires than braking action Experimentally, it has been shown (Ref 12) that tire wear varies exponentially, rather than linearly, with force This is general agreement with Eq 2 and 4

The above models explain regular wear in tires reasonably well However, tire also exhibit irregular wear, and at present there are no good mathematical models to explain this phenomenon According to Livingston (Ref 20), irregular wear may occur because of differences in stiffness and ground pressure from point to point in the tire footprint area These differences result in different strain levels within the contact area, leading to different amounts of sliding from point to point Livingston has proposed a theoretical approach that assumes differential slips for various tread elements to explain irregular wear He suggests using finite element analysis (FEA) simulation in incremental steps

Standard Wear Rate

Tire design and construction parameters affect the nature of the forces generated in the footprint area As mentioned earlier, force is the dominant controlling factor in tire wear Therefore, to compare wear rates for different tire constructions, the measured wear rate for each construction can be related to a reference force value and corrected for surface temperature variation This corrected rate is known as the standard wear rate Incorporating Schallamach's temperature-correction relationship (Ref 8), Veith (Ref 12) developed the following equation for the standard wear rate:

(Eq 5)

where W°50 is the standard wear rate at 50 °C (120 °F) and 100 lb cornering force, F is the average cornering force in pounds, is the temperature coefficient of wear rate, and T is the test temperature in degrees Celsius

Fleet Testing and Trailer Testing

Measurement of tire wear is generally done by fleet testing A car fitted with one type of tire is driven over specific routes representative of typical driving conditions Test conditions, such as tire rotation, frequency of vibration, speed, straight driving, turns, and nature of pavement, are preset The initial and final tread depths averaged over all the grooves are measured, and the total wear is calculated from the difference between the tread depth values Based on the miles driven,

an "average" wear rate is obtained The test is repeated using tires with different variations in design, compound, materials, and so on Results allow comparison of the wear rates of different tires The primary limitation of this method

is that the various factors that influence tread wear cannot be independently controlled

In the trailer testing method, both cornering force and slip angle can be controlled and monitored This method allows the researcher to directly control the magnitude of the force acting on the tire Thus, wear rate as a function of force can be determined Veith (Ref 12) describes a trailer test system with elaborate instrumentation for controlling test parameters and for making measurements He used this method to study the effects of tire force, weather conditions, and pavement

texture on wear rate He also showed that wear rate is proportional to F2, where F is force, which is in agreement with Eq

2 Driving severity number (DSN) is a concept developed by Veith (Ref 21) to arrive at a composite parameter value that combines the total effect of the three factors mentioned above Using the DSN, Veith (Ref 21) quantified the tire input

force and driver vehicle interaction over a test course to better explain tread wear rankings

Effects of Load and Velocity

Veith (Ref 12) has shown that lateral force, F, can be expressed in terms of vehicle weight, L, as:

(Eq 6)

where V is vehicle velocity, R is the radius of curvature, and G is acceleration due to gravity L is equal to four times the

average tire load Combining Eq 6 and Eq 2, the wear rate can be written as:

(Eq 7)

where K is a constant This indicates that the wear rate is proportional to the second power of tire load and to the fourth

power of vehicle velocity

Experimental verification of Eq 7 is not straightforward It has been found that the extent of change in W with change in L depends on how W is measured by trailer method or by vehicle cornering For the same change in L, the corresponding change in cornering force is much smaller in the former test procedure Therefore, change in W is higher in the vehicle

cornering method than in the trailer method

Again, the relationship between wear rate and vehicle velocity also seems to be method dependent Veith (Ref 12) found that the effect of vehicle speed was very small in trailer testing and had a more pronounced effect in fleet testing

Experimentally, it was found that the wear rate, W, was approximately proportional to the cube of the velocity (Ref 4) It

appears that there is some inconsistency between theory and experiment, as well as between different experiments, and it

is thus difficult to correlate experiment with theory

Test Severity

The major components of test severity are tire force severity, pavement surface severity, and weather/temperature severity Biard and Svetlik (Ref 22) reported that the wear index of styrene-butadiene rubber ranged from 98 to 216 (natural rubber = 100) as severity increased Geesink and Prat (Ref 23) found that the wear rating order of different tread compounds changed with test severity Veith (Ref 12, 13) systematically studied tread wear as a function of the three severity factors listed above

Tire Force Severity. Force severity was controlled by different vehicle cornering and braking maneuvers Miller et al

(Ref 24) have shown that the cornering wear rate is about seven times higher than the braking wear rate These results confirm the previous assertion that cornering force is the most important variable affecting tire wear

Veith (Ref 12, 13) varied tread composition by using different types of carbon black and determined the tread index as a function of cornering force Severity index gradient was defined as the rate of change of wear index with force severity This experiment led to the conclusion that carbon black of smaller particle size provides good wear resistance at high severity conditions Studebaker (Ref 25) and Dannenburg and Amon (Ref 26) also support this finding As radial tires became popular, the wear rating of radial construction versus bias construction was measured It was found that as force severity increases, the radial construction has a higher wear index than the bias construction In other words, radial tires have higher wear resistance than bias tires

Pavement surface severity is a combination of micro- and macrotexture and is qualitatively ranked as harsh or blunt Lowne (Ref 27) has shown that pavement texture is important in determining wear rate The wear ranking of tread compounds is not independent of the surface texture of the pavement For example, a compound that is superior to another compound on a blunt surface can become inferior on a harsh surface Veith (Ref 13) observed similar ranking reversals in testing of two different tread compounds on blunt and harsh surfaces This sort of ranking reversal has also been observed in laboratory abrasion experiments by Schallamach (Ref 28), Kragelski and Nepomnyashchi (Ref 29), and Pulford (Ref 30) Schallamach explains this behavior by invoking the tensile tear mechanism and/or physical property degradation of rubber, depending on the nature of the abrading surface

Weather Severity. Tire surface temperature, ambient temperature, rain, and snow affect wear Wear rate increases as

temperature increases Veith has studied this phenomenon extensively (Ref 13, 16) The literature presents conflicting results with regard to wear rate on wet versus dry pavement

Effects of Various Factors: Experimental Design Study

Veith (Ref 14) conducted and extensive study of the effects of multiple levels of five design factors on tread wear (Table 2) The design details, the tread wear testing, and the regression analysis are presented in Ref 14 The factors in order of importance for superior wear resistance are tire type, tread loss modulus, and tread pattern Radial tires had better wear resistance Similarly, as the loss modulus increased, wear resistance increased Tires with higher groove volume and aspect ratio showed poor wear resistance

Table 2 Effects of tire-related factors on tire wear rate

Tire type (ply)

Tread loss modulus

1.37 MPa (200 psi)

0.84 MPa (122 psi) 1.37 MPa (200 psi) -15

(a) Minus sign represents an improvement in wear rate; plus sign represents a decline in

wear rate:

Veith and Chirico (Ref 13) also studied carbon black in a separate experiment At high severities, carbon blacks with high structure and surface area exhibited better wear resistance than those with normal structure The rate of wear passed through a minimum as a function of carbon black level at each severity Wear rate increased with increased oil content Some crossover results were obtained as the severity level changed from high to low

Tire Conceptual Model

A tire is a composite deformable structure Under load, each tire component deforms The magnitude of these deformations varies from component to component, depending on component stiffness The different deformations generated in each component cause interfacial stresses, leading to a complex stress distribution in the footprint area Oblizajek and Veith (Ref 15) discussed a detailed experimental setup to measure the footprint stress distribution and mechanical properties of tire components

The mechanical properties discussed in Ref 15 can be used to characterize wear using the Gough model (Ref 11, 31) The Gough model considers the tire tread band as a beam and the carcass as an elastic foundation Tread/belt transverse

flexural rigidity EI, and carcass foundation stiffness, K, are introduced in this model The flexural rigidity is a product of Young's modulus, E, and the moment of inertia, I The effects of different combinations of EI and K values are discussed

in Ref 11 It is reported that tire wear improves as the pairs of EI and K values change in the following order: high K, low EI; high K, high EI; low K, high EI Using composite theory, Walter et al (Ref 32) developed a method to calculate the

tire component stiffness parameters in different directions

Further refinement of the Gough model introduced tread band transverse shear modulus, G, as an additional important parameter to explain wear resistance The modified tread band transverse stiffness incorporates both EI and G (Ref 31)

Oblizajek and Veith (Ref 15) have shown very good correlation between these mechanical properties and tread wear over

a wide range of tire construction

Coefficient of Friction of Rubber. Providing a specific value for the friction coefficient of rubber is impossible because there is no one value for the coefficient of friction of rubber against a specific surface (asphalt, concrete, glass,

and so on) Friction is dependent on a complex system of parameters (that is, normal force, sliding and rotating speed, surface roughness, surface lubrication, and temperature) Grosch (Ref 33) has conducted extensive experimental studies

of rubber friction over a wide range of parameters and conditions It is postulated that the total frictional force, Ft, is given

by the equation:

where Fa, Fh, and Fv are the adhesive, the hysteretic, and the viscous components, respectively In addition, the magnitude

of the friction force depends on the contact area Muhr and Roberts (Ref 34) have published a review of rubber friction and its relation to wear Despite all of the experimental work done or rubber friction to extrapolate data from the laboratory results, determining the actual tire tread wear rate in the field is still not easily possible

Future Outlook

Tire tread wear is a result of dissipation of frictional work generated in the footprint area Cornering force is the primary factor responsible for tire wear Reasonable correlation between tire component mechanical properties and tread wear has been established by statistical methods However, this approach has its limitations Although much research has been directed toward gaining an understanding of tire wear, most of it so far has been empirical More research is needed to establish a precise relationship among composite properties, material properties, and tire wear

References

1 A Schallamach, Rubber Chem Technol., Vol 41, 1968, p 209

2 K.A Grosch, Rubber Chem Technol., Vol 65, 1992, p 78

3 K.A Grosch, Fractography of Rubber, A.K Bhowmick and S.K De, Ed., Elsevier, 1991, p 139

4 C.T.R Pulford, Rubber Chem Technol., Vol 58, 1985, p 653

5 A.H Muhr and A.D Roberts, Paper presented at the Rubber Division ACS Meeting (Toronto), 21-24 May

1991

6 A Schallamach, Trans Inst Rubber Ind., Vol 28, 1952, p 256; J Polym Sci., Vol 9, 1952, p 385

7 K.A Grosch and A Schallamach, Trans Inst Rubber Ind., Vol 41, 1965, p 81

8 A Schallamach and K.A Grosch, Wear, Vol 4, 1961, p 356

9 A.N Gent, and C.T.R Pulford, J Appl Polym Sci., Vol 28, 1983, p 943

10 V.E Gough, Trans Inst Rubber Ind., Vol 32, 1956, p 27

11 V.E Gough, Wear, Vol 2, 1958, p 107; Paper 667A, presented at SAE Meeting, Detroit, 19-21 Mar 1963

12 A.G Veith, Rubber Chem Technol., Vol 46, 1973, p 801, 821

13 A.G Veith and V.E Chirico, Rubber Chem Technol., Vol 52, 1979, p 748

14 A.G Veith, Tire Sci Technol., Vol 14, 1986, p 201, 219, 235

15 K.L Oblizajek and A.G Veith, Tire Sci Technol., Vol 14, 1986, p 264

16 A.G Veith, Paper presented at the Rubber Division ACS Meeting (Toronto), 21-24 May 1991

17 "Vehicle Dynamic's Terminology," J670e, 1992 SAE Handbook, Vol 4, On-Highway Vehicles and

Off-Highway Machinery, Society of Automotive Engineers, p 34.246

18 A Schallamach and D.M Turner, Wear, Vol 3, 1960, p 1

19 D.I Livingston, Paper presented at the Rubber Division ACS Meeting (Mexico City), May 1989

20 D.I Livingston, Plenary Lecture, Tire Society Meeting, University of Akron, 15 Mar 1991

21 A.G Veith, Tire Sci Technol., Vol 14, 1986, p 139

22 J Biard and F Svetlik, Rubber Chem Technol., Vol 26, 1953, p 731

23 H.A.O.W Geesink and L Prat, Rubber Chem Technol., Vol 31, 1958, p 166

24 R.F Miller, R Marlowe, and J.L Ginn, Rubber Plast Age, Vol 42, 1961, p 968

25 M.L Studebaker, Rubber Chem Technol., Vol 41, 1968, p 373

26 E.M Dannenburg and F.H Amon, Rubber World, Vol 131, 1955, p 627

27 R.W Lowne, Rubber Chem Technol., Vol 44, 1971, p 1159

28 A Schallamach, Rubber Chem Technol., Vol 31, 1968, p 209

29 I.V Kragelski and E.F Nepomnyashchi, Wear, Vol 8, 1965, p 303

30 C.T.R Pulford, J Appl Polym Sci., Vol 28, 1983, p 709

31 V.E Gough, Kautsch Gummi, Vol 20, 1967, p 469

32 J.D Walter, G.N Augeropoulos, M.L Janssen, and G.R Potts, Tire Sci Technol., Vol 1, 1973, p 210

33 A Grosch, The Speed and Temperature Dependence of Rubber Friction and Its Bearing on the Skid

Resistance of Tires, The Physics of Tire Traction, D.F Hays and A.L Browne, Ed., Plenum Press, 1974, p

143

34 A.H Muhr and A.D Roberts, Friction and Wear, Natural Rubber Science and Technology, A.D Roberts,

Ed., Oxford University Press, Oxford, 1988, p 773

Friction and Wear of Aircraft Brakes

E.M Tatarzycki and R.T Webb, Aircraft Braking Systems

Introduction

AIRCRAFT BRAKES are designed to stop an aircraft through the conversion of kinetic energy into heat by the mechanism of friction Aircraft brakes are composed of multiple disk pairs, which are commonly referred to as the brake heat sink Unlike an automotive brake, which consists of a single piston, a single rotor, and two brake pads, the majority

of aircraft brakes use full-circle rotors and stators (Fig 1) Brake housings normally contain several pistons for applications of the normal force needed to develop the brake torque Large commercial transport aircraft brakes must be capable of absorbing up to 135 MJ (100 × 106 ft · lbf) of energy The high levels of torque developed to stop an aircraft require the conversion of large amounts of kinetic energy into thermal energy over a short period of time This energy conversion process produces very high energy fluxes at the multiple friction interfaces, resulting in high temperatures and stresses in the brake heat sink

Fig 1 Illustrations of steel and carbon aircraft brakes

The friction and wear characteristics of the friction materials used in aircraft brakes are influenced by internal factors (such as friction-material composition and heat sink mass) and external factors (such as the amount of kinetic energy absorbed by the brake, the surface velocity of the friction interfaces, and aircraft deceleration requirements) Simply put, these factors control the temperature at the interface and the normal and tangential forces of the friction material

Aircraft Friction Materials

Three basic friction materials are presently being used in the aircraft brake These include:

• Sintered metal friction materials

• Carbon-carbon composites

• Organic materials

The sintered metallic friction material is probably the most widely used in aircraft brakes The steel brake heat sink consists of a sintered metallic friction material bonded to a steel supporting backing plate The steel brake has a higher wear rate, is heavier, and has a lower cost per brake landing compared to the carbon brake The carbon/carbon composite friction material is the latest technology The carbon brake is lighter in weight, has excellent high-temperature performance, has a low wear rate, and has a higher cost per brake landing Organic brake linings are used on older and

smaller aircraft and in helicopter rotor brakes Most of the work being done in this area is aimed at replacing the existing asbestos-based linings with nonasbestos materials and will not be discussed in this article

Steel Brakes

The metallic friction materials are primarily fabricated into two final forms The current designs dictate that the metallic friction material be bonded into metal cups These linings, as they are commonly called, are then riveted to a carrier disk The other form is called a full-sintered disk on which the friction material is bonded directly to the carrier disk

The steel brake is composed of multiple pairs of rotors and stators (Fig 1) In this type of brake, the stators carry the metallic friction material and the rotors are composed of a high-strength high-temperature steel, commonly called the opposing surface Some designs have the rotors carrying the metallic friction material The opposing steel typically will last two to three times longer than the metallic friction material lining The selection of the metallic friction material influences the brake design and must be carefully considered in order to obtain optimum dynamic performance, friction coefficient, stability, and wear rate of the friction pair The system dynamics, which includes brake, wheel, tire, strut, and hydraulics, is also greatly influenced by the selection of the friction material

Chemistry. The metallic friction materials are blends of various metallic and nonmetallic powders The mixtures are primarily blends of copper and iron with one or the other being predominate in the composition The base matrix of either iron or copper is modified by the addition of graphite, in natural or synthetic form; silicon as an abrasive; and modifiers such as high-temperature lubricants (for example, molybdenum disulfide) The size and shape of the powders can influence the final product Proprietary materials may be added to control brake vibration, matrix strength, and wear

Processing. The metallic mixture is then sintered in a furnace to form a metal-matrix composite The sintering process takes place at high temperatures and sometimes under pressure The required final density of the metallic matrix is attained by hot or cold working of the lining The choices made in the processing of this matrix can greatly affect the dynamic performance and wear life of the friction pair The bond of the friction material to the steel cup or core is critical

If the sintered matrix does not bond, then the matrix can spall and leave large areas void of friction material The bonding

of the matrix to cup or core is controlled by the base metallic elements (iron and copper) and any devices that might be added to the design to enhance matrix-to-carrier bond strength

The copper-base metal matrix can attain a good bond to the carrier if sintered at an optimum temperature, time, and pressure The iron-base metal matrix sometimes requires a screen welded to the cup bottom to supplement bonding The iron-base metal matrix relies on a mechanical bond occurring in the open area of the screen

Friction Material Selection. Choice of friction material determines the friction coefficient and the brake design The iron-base matrix tends to have a lower coefficient This requires the brake to have large-diameter pistons to achieve specific design goals (such as stopping distance, which is a critical parameter in aircraft certification) The copper-base matrix tends to have a higher coefficient This would require smaller pistons and result in a lighter housing

Thermal Properties. The thermal and strength properties are influenced by the chemistry, particle shape, and particle size of the friction material mixture For a stator-rotor friction pair, the most significant parameter is the temperature at the friction interface Typically, the lower the interface temperature, the better the brake wear life This temperature is influenced by the thermal conductivity of the lining and opposing steel surface, the thickness of each disk, and other boundary conditions It is desirable to design a friction pair such that the heat is conducted away from the interface quickly

Balance. The correct distribution of mass among the various brake members is crucial for optimum dynamic performance and balanced disk wear The specific distribution of mass in the brake is very dependent on the nature of the friction mix Each interface of a multiple-pair brake does not perform equally Typically, the greatest wear takes place on surfaces that are nearest the pistons Therefore, the friction pairs are normally of greater axial thickness closer to the pistons Mass balancing of the brake stack components is done so that the friction interfaces all wear out at the same time

Brake wear is normally measured by the length of a steel pin (wear pin) extending from the pressure plate through the housing If a brake has a 25 mm (1 in.) wear pin, this means that 25 mm (1 in.) of wearable material is left in the brake stack Brake wear is usually expressed in units of mm/face/stop In subscale dynamometer evaluations, weight-loss measurements are of primary importance

The wear mechanism taking place at the friction surface is very complex and not fully understood After the first few stops, both surfaces form a film, sometimes referred to as a layer of glaze, at the contact surface This film is mainly an oxide, the mechanical and tribological properties of which are different than the parent materials The film thickness varies over the surface The combination of thickness and properties controls the resulting coefficient and wear rate of the brake

Wear Mechanism. An aircraft brake is expected to operate over a wide range of energy inputs These varying energies control the temperature at the friction interface During low energy, that is, low interface temperature, the wear is characterized by abrasive wear Abrasive wear occurs when hard particles cut grooves across the opposing surface and displace material When brake energies are higher, the interface temperatures are also higher and an adhesive type of wear takes place Adhesive wear occurs when surface asperities, bonded together under localized high temperatures and pressures, shear apart during sliding of the surfaces The wear debris may become trapped between the surface and contribute to abrasive wear

Wear Rates. Figure 2 shows wear rates based on normalized stack loading and average energy flux (AEF) For an aircraft to attain 1000 stops (a stop is equivalent to a landing), assuming a five-pair brake with a 50 mm (2 in.) wear pin, wear rates of 0.0025 mm/face/stop (0.0001 in./face/stop) would be required

Fig 2 Typical wear rates for a steel brake based on stack loading and average energy flux (AEF)

When aircraft taxi out to the runways, they normally start and stop along the way These small stops are referred to as taxi stops and snubs During a snub an aircraft does not come to a complete stop Brake life typically is measured by the number of landings, and this would include any taxis and snubs required to maneuver the aircraft For steel brakes, the wear that takes place at these low energies is not as significant as the main landing wear and is normally considered as being part of the main landing stop In contrast, the taxi wear in carbon brakes accounts for a large portion of the brake wear

Aircraft brakes are designed to handle one rejected take-off (RTO) stop (the wear rate is normally one hundred to one thousand times greater during an RTO stop than during a service stop) After an RTO stop, the brake and wheel are normally scrapped In recent years the worn RTO has become an important issue Because RTO wear on a worn brake stack in normally at a higher rate than wear on a new stack, some aircraft have been mandated by the Federal Aviation Administration (FAA) to limit the brake life to ensure that there is sufficient wear material in the brake to perform a worn RTO

Brake Design. The selection of the metallic mix can influence the brake design in many ways A friction material with

a low friction coefficient will require larger pistons to supply a large clamping force (normal force) in order to achieve the high torque required to attain stopping distances The required heat-sink mass is normally determined by how much energy the brake must absorb during the RTO stop or by some other critical condition Typically, new brakes with metallic linings are designed to be loaded in the 750,000 to 900,000 J/kg (250,000 to 300,000 ft · lbf/lb) range for the RTO stop The required aircraft deceleration is also influenced by mix selection The deceleration rate controls the

average energy flux at the interfaces, and different types of mixes have differing limitations The limitations on AEF determine the amount of rubbed area required in the brake

Coefficient of Friction. In general, metallic friction materials exhibit a consistent coefficient of friction over a broad range of loadings, whereas carbon friction materials exhibit a wider band of coefficient over the operating spectrum of the brake The balance coefficient is influenced by the past service history of the friction material, the loadings, and the design of the opposing surface Normally, the brake coefficient decays during the first several landings when starting with new friction material This decrease in coefficient is usually attributed to formation of oxides and stabilization of the surface film

The friction coefficient is also affected by the energy flux and the total loading on the brake (Fig 3) If a brake design requires a high energy flux, the coefficient must then be accounted for Normally, as the energy flux is increased, the average coefficient decreases, and thus the brake clamping forces must be of sufficient magnitude to ensure that the required deceleration rate is met

Fig 3 Typical range of friction coefficients for a steel brake based on stack loading

The brake loading is another parameter that changes the average coefficient during a stop The requirements of the RTO stop normally influence the brake design, and lower friction coefficients due to higher loading and surface velocity must

be considered Coefficients are also influenced by the material and design of the opposing metal surface The majority of the steel brakes use AMS 6385C (plate), AMS 6302C (forging), or Timken 17-22AS for the opposing surface

Spalling. The metallic friction material is exposed to high stresses and temperatures During the extended use of an aircraft brake, the linings wear thinner and the friction material oxidizes; both factors potentially lower the matrix strength

of the friction material The high stresses encountered during stopping may spall all the friction material The loss of large chunks of friction material reduces the rubbing contact area, thereby reducing brake torque Once a lining spalls, the metallic cup that holds the friction material can warp In severe cases, the rivet could fail and deline the face of one disk

In order to eliminate the potential for matrix spalling, minimum lining thicknesses have been established and friction materials with enhanced matrix strength have been developed

Vibration. Typically, a metallic friction material that has a longer wear life also tends to cause brake vibration It has been determined that a friction material with an increasing coefficient during a stop will tend to make the braking system unstable and cause vibration (Fig 4) Proprietary additives and changes in the basic constituents of the friction material are made to produce a "smooth" friction material Severe brake vibrations can cause the strut to vibrate, possibly leading

to strut failure; therefore, such vibrations must be accounted for and remedied

Fig 4 Variation of friction coefficient versus tire speed during a single stop

Service Life. The aircraft industry is constantly striving to produce longer wearing brake materials Typically, large commercial jets get around 1000 landings on a brake stack before refurbishment The individual brake components should

be balanced so that they wear evenly Uneven wear can cause problems, such as spalling and delining

Carbon Brakes

Raw Materials. Carbon-carbon friction materials are composites comprised of high-density carbon fibers embedded in

a carbon matrix At the present time, all the carbon fibers used in carbon aircraft brakes are made from two precursors: polyacrylonitrile (PAN) or pitch Fiber properties are normally controlled by the manufacturing process of the fiber and are beyond the scope of this article Such information, however, can be found in the articles “Carbon Fibers” and

“Processing of Carbon-Carbon Composites” in Composites,Volume 21 of ASM Handbook

Carbon fibers can be woven, knit, spun, and generally handled like most textile threads, but with more difficulty In brakes, woven fabrics, short-length yarns, chopped fabrics, and woven three-dimensional preforms are used Each manufacturer has its own preferences; thus the entire system reinforcement and matrix can be designed to produce specific performance

The carbon matrix can be established in many ways The most common is by the deposition of pyrolytic carbon directly onto the surfaces of the pores and voids within the disk This technique is called the chemical vapor infiltration (CVI) or chemical vapor deposition (CVD) process The infiltration or deposition is carried out in a reactor (furnace) under vacuum, high temperature, and the flow of a hydrocarbon gas Another way to achieve a carbon matrix is to mold the fibers with a resin, such as phenolic, and convert the resin to carbon by charring at temperatures ranging from 540 to 815

°C (1000 to 1500 °F) Subsequent impregnation as with resin, pitch, or resin-pitch blends followed by a char cycle can also be carried out

Processing. Carbon-carbon composites are porous because of the nature of the densification process The conversion of phenolic resin to char results in porous, glassy carbon The CVI process relies on the ability of the gas to infiltrate the maze of fibers or partially densified composite to deposit a carbon, which some researchers have categorized into three groups: isotropic, smooth laminar, or rough laminar As the infiltration proceeds, many of the interior surfaces become inaccessible to the gas These remain inaccessible and become permanent voids Consequently, a multiple-step densification cycle is necessary to reduce the porosity and to achieve acceptable densities, usually in the 1.70 to 1.85 g/cm3 range

Processing may also include heat treatment at temperatures ranging from 1650 to 2800 °C (3000 to 5000 °F) Such a heat treatment is called graphitization and is used to control the properties of the carbon For example, a fully graphitized carbon (graphite) will have a crystalline form, have an atomic lattice spacing of 0.3359 nm (3.359 ), offer little resistance to indentation (making it soft), have a high coefficient of friction, and will exhibit high wear An ungraphitized carbon will be amorphous in nature, have an atomic lattice spacing of 0.344 nm (3.44 ), show high resistance to indentation (making it hard), produce a low coefficient of friction, and exhibit low wear By selecting the appropriate raw material constituents, quantity of each constituent, specific process, and heat treatment temperature, it is possible to engineer a carbon for a range of friction coefficients and wear rates

Physical Mechanical, and Thermal Properties. Carbon-carbon composites used in brake applications are unique because a single material serves as the friction surface, the heat sink, and the structural member of the brake Consequently, this family of materials must have a wide range of properties Densities must be high enough to enable the brake to absorb the high kinetic energies expended by the largest transports, yet low enough to be considered as a light-weight aircraft material Its surfaces must produce a suitable rubbing aggressiveness to stop a large aircraft, yet maintain uniformity in friction over a wide range of temperatures As a structural member, the carbon must exhibit high strength at both room and elevated temperatures

The following properties are considered to be important for carbon brake performance:

The requirements for the RTO normally determine the adequacy of the brake design A sufficiently high coefficient is needed to stop the aircraft Proper selection of raw materials and processing conditions will determine the coefficient under RTO conditions as well as under the other energy conditions Normally, the highest coefficient occurs at taxi conditions; the lowest occur at RTO energies Figure 5 shows a typical variation in coefficient from taxi to RTO energies

Fig 5 Variation in coefficient of friction of carbon brake from taxi (left side) to RTO (right side) energies

Brake Wear. Carbon aircraft brake wear characteristics depend upon the raw materials and the specific processing parameters used during fabrication as well as the operating conditions of temperature, rubbing velocity, contact pressure, and prior braking history Published information on the wear of carbon brakes is sparse, primarily because of the highly competitive nature of the aircraft brake business Consequently, brake wear will be discussed in terms of possible underlying mechanisms rather than published scientific data

Aircraft brakes operate under a wide range of energy conditions Typically, an aircraft will taxi out to the runway, making several braking snubs (slowdowns) and complete stops along the way After a successful take-off and flight, the landing occurs Landings are classified into three typical categories:

Fig 6 Variation in wear of carbon brake from taxi to RTO energies

Each of the three energy ranges has its own wear mechanism The taxi region exhibits the highest wear, and mechanical abrasion predominates In the mid-energy (service energy) region, wear is lowest because a thin film protects the substrate The third region occurs when temperatures are sufficiently high to produce oxidation of any debris or substrate (fiber or matrix) Oxidation usually occurs during overload and RTO conditions

PAN-base fibers wear differently than pitch-base fibers; these differences are associated with fiber processing as well as composite processing A harder matrix will wear less than a softer one Studies have shown that circumferentially oriented fibers have high wear and a high coefficient of friction

Wear Mechanism and Wear Debris Analysis. The worn surfaces of the carbon materials have bands of different reflectance when observed via macro examination Debris material occurs in both bright and dull bands The observed banding on the wear surface of the composite is related to variations in reflectance; these variations are thought to be related to differences in the character of the wear debris The bright bands consist of a thin film of debris with a polished appearance that produces a high reflectance The dull bands have a high number of fissures or shallow grooves and scratches in the longitudinal fiber bundles, both of which are indicative of fiber removal caused by abrasion The debris material in the dull bands does not produce a high reflectance and appears less dense (more porous) than the bright bands The debris in the dull bands is particulate and is comprised of fibers and matrix This debris does not form a film, but fills

in the original porosity, mainly in the matrix, and is found primarily under low-energy conditions

The debris in the bright band consists of a thin film (up to a few micrometers thick) that is smeared over fibers and matrix and also fills in the pores The bright bands appear to be denser (less porous) than the dull bands The wear debris film is grooved, whereas the fibers and matrix are not This indicates that the wear debris film is protecting the fibers and matrix from abrasion

The thin film of wear debris is amorphous in character and is comprised of both fibers and matrix High-energy braking conditions produce a high coverage of wear debris (film), which may result in a low coefficient of friction and low wear rate More detailed information on wear debris analysis can be found in the article "Lubricant Analysis" in this Volume

Moisture Problems. Carbons and graphites have an affinity for moisture These materials adsorb moisture; that is, water molecules are attracted to their exterior and interior surfaces Consequently, the presence of moisture significantly reduces the coefficient of friction When a brake is sitting for a period of several hours, the rubbing surfaces will adsorb moisture from the air Thus, when a brake stop is initiated in this condition, the stopping power is significantly reduced because of the low coefficient This is typically referred to as "morning sickness." As rubbing continues through the stop, the moisture evaporates and the coefficient returns to its normal dry value

Oxidation. Carbons and graphites are also subject to oxidation at elevated temperatures Typically, the threshold of

oxidation is considered to be 430 °C (805 °F) Technically there is oxidation at this temperature, but it is so low that it is considered to be negligible Other than taxi conditions, the operating temperatures of a carbon brake will range between

500 and 1000 °C (930 and 1830 °F) during landings; under RTO energies the temperatures exceed 1300 °C (2370 °F) and oxidation becomes a significant factor In order to keep the disks from deteriorating, an oxidation inhibitor is applied to the outer and inner diameters of each carbon brake disk Inhibitor is not applied to the wear surfaces because it would alter the friction characteristics of the carbon

Friction Coefficient Variability. As stated previously, friction coefficients have a larger range over the operational spectrum of the aircraft when compared to the friction coefficients for steel brakes This coefficient range has implications for brake control (antiskid) system design In the carbon friction material, friction coefficients can vary by factors of 3 or more over the operation range of the brake Therefore, the brake torque can vary by a factor of 3 or more Carbon processing and brake frame design must take this friction coefficient variation into account

Vibration. An aircraft wheel and brake is a system with multiple degrees of freedom that is subjected to high dynamic loads This high dynamic loading is transient in nature and can be an exciter of vibration

In a carbon brake, the most critical vibration mode is known as whirl This mode consists of an accordian-type action of the brake disks combined with an orbiting motion of the brake structure about the axle (Fig 7) The torque output and friction coefficient of the brake, as functions of velocity and brake pressure, are significant parameters in determining whether or not whirl motion will occur and, if it does occur, the severity of the vibration Severe whirl can result in damage to the carbon brake disks and other wheel/brake hardware

Fig 7 Carbon brake vibration

There are two basic approaches that can be taken to control whirl vibration First, carbon friction coefficient characteristics are engineered in such a way that the brake coefficient, at a given velocity and brake pressure, remains low enough to minimize the potential for the whirl vibration Second, the aircraft wheel/brake is designed to incorporate the required stiffnesses and damping characteristics Adequate stiffnesses and damping, combined with favorable friction coefficient characteristics, will ensure the stability of the aircraft wheel/brake structure at given dynamic conditions

Testing

Friction material testing is conducted on either direct-connected dynamometers or on landing wheel dynamometers Typically, new friction materials are screened in subscale brakes on direct-connected dynamometers Once a potential material has been selected, then full-scale brakes are tested on landing wheel dynamometers which use an aircraft tire and wheel Full-scale brake testing is very expensive and therefore limited During aircraft wheel and brake qualification testing, the full-scale brake is run through numerous tests before any aircraft testing is done

Test Requirements. Generally an aircraft brake must pass testing standards set up by the military or FAA The military requirements are outlined in MIL-W-5013, and the FAA requirements are continued in TSO-C26 The Society of Automotive Engineers (SAE) also publishes an Aerospace Recommended Practice (ARP) for aircraft wheels and brakes (ARP 597) The airframe manufacturers also specify extensive supplemental qualification requirements that must be met before the brake can be qualified for service

Selected References

• J.F Archard, The Temperature of Rubbing Surfaces, Wear, Vol 2, 1958/59

• F.P Bowden and J.E Young, Proc R Soc (London) A, Vol 208, 1951, p 444

• H.W Chang, Correlation of Wear with Oxidation of Carbon-Carbon Composites, International

Conference on Wear of Materials, American Society of Mechanical Engineers, 30 Mar to 1 Apr 1981

• T.S Eyre and F Wilson, Wear of Grey Cast Iron Under Unlubricated Sliding Conditions,

ASME/ASLE International Lubrication Conference (New York), 9-12 Oct 1972

• D.B Fischbach and D.R Uptegrove, Oxidation Behavior of Some Carbon/Carbon Composites, 13th

Biennial Conference on Carbon (Irvine, CA), 1977

• K Gopinath, G.V.N Rayudu, and R.G Narayanamurthi, Friction and Wear of Sintered Iron, Wear,

Vol 42, 1977, p 245-250

• B Granoff, H.O Pierson, and D.M Schuster, Carbon-Felt, Carbon-Matrix Composites: Dependence

of Thermal and Mechanical Properties on Fiber Volume Percent, J Compos Mater., Vol 7, Jan 1973

• T.-L Ho, "Development and Evaluation of High-Energy Brake Materials," Ph.D thesis, Rensselaer

Polytechnic Institute, 1974

• J.M Hutcheon and M.S.T Price, The Dependence of the Properties of Graphite on Porosity,

Proceedings of the Fourth Conference on Carbon, Pergamon Press, 1960

• W.V Kotlensky and P.L Walker, Jr., Crystallographic and Physical Changes of Some Carbons Upon

Oxidation and Heat Treatment, Proceedings of the Fourth Conference on Carbon, Pergamon Press,

1960

• I.V Kragelskii, Friction and Wear, Butterworths, Washington, 1965, p 117

• J.K Lancaster, Instabilities in the Frictional Behavior of Carbons and Graphites, Wear, Vol 34, 1975

• R.L Lewis and R.E Raymond, "Stopping Distance Analysis," Society of Automotive Engineers, Inc.,

Paper No 730193, 1973

• F.F Ling and E Saibel, On Kinetic Friction Between Unlubricated Metallic Surfaces, Wear, Vol 1,

1957/58

• J Molgaard and V.K Srivastava, The Activation Energy of Oxidation in Wear, Wear, Vol 41, 1977

• N Murdie, C.P Ju, J Don, and F.A Fortunato, Microstructure of Worn Pitch/Resin/CVI C-C

Composites, Carbon, Vol 29, 1991, p 335-342

• D Pavelescu and M Musat, Some Relations for Determining the Wear of Composite Brake Materials,

Wear, Vol 27, 1974

• T.F.J Quinn, A.R Baig, C.A Hogarth, and H Muller, Transitions in the Friction Coefficients, the

Wear Rates, and the Compositions of the Wear Debris Produced in the Unlubricated Sliding of

Chromium Steels, ASME/ASLE International Lubrication Conference (New York), 9-12 Oct 1972

• E, Rabinowicz, Friction and Wear of Materials, John Wiley & Sons, 1965

• D.M Rowson, The Interfacial Surface Temperature of a Disk Brake, Wear, Vol 47, 1978

• L Rozeanu, Friction Transients (Their Role in Friction Failures), Trans ASLE, Vol 16, 1975, p

257-266

• J.J Santini and F.E Kennedy, Jr., An Experimental Investigation of Surface Temperatures and Wear

in Disk Brakes, Lubr Eng., Aug 1975

• P Stanek, N Murdie, E.J Hippo, and B Howdyshell, The Effect of Fiber Orientation on Friction and

Wear of C-C Composites (Extended Abstracts), Biannual Conference on Carbon (Santa Barbara, CA),

1991, p 378-379

• I.L Stimson and R Fisher, Design and Engineering of Carbon Brakes, Philos Trans R Soc

(London) A, Vol 294, 1980

• E.M Tatarzycki, Friction Characteristics of Some Graphites and Carbon Composites Sliding Against

Themselves, 13th Biennial Conference on Carbon, 1977

• A.K Vijh, The Influence of Solid State Cohesion of Metals and Non-Metals on the Magnitude of

Their Abrasive Wear Resistance, Wear, Vol 35, 1975

Wear of Jet Engine Components

J.D Schell and K.P Taylor, General Electric Aircraft Engines

Introduction

A JET ENGINE is a sophisticated piece of machinery with many moving parts; the potential for wear problems exists whenever moving parts come into contact or unintended motion occurs between stationary contacting parts As for any system, wear in jet engines can be controlled through proper design, material selection, and lubrication A schematic cross section of a typical jet engine is shown in Fig 1 The major engine subsystems consist of the fan, the high-pressure compressor (HPC), the combustor, the high- and low-pressure turbines (HPT and LPT), and the exhaust nozzle The engine design contains one nonrotating system and two concentric rotating systems The nonrotating (stator) system is made up of structural frames and casings The low-pressure rotating system consists of the fan disk(s) and fan blades, the LPT disks and turbine blades, and a connecting shaft The high-pressure rotating system consists of the HPC disks/spools and compressor blades, the HPT disks and turbine blades, and a connecting shaft

Fig 1 Jet engine cross section showing important subsystems and potential areas of wear

Operating environments vary widely between different sections of the engine and depend on where the engine is in its mission Temperatures may vary from subzero to above 1095 °C (2000 °F), rotational speeds may climb to more than 15,000 rev/min, and contact loads may range from a few psi to local hertzian stresses well beyond 1720 MPa (250,000 psi) in rolling-element bearings The relative motions of components may be unidirectional sliding of rotating parts on stators, oscillatory sliding varying from a few thousandths of an inch up to several tenths of an inch, or vibratory motion resulting in impact between components Components also may be subjected to ingested particle impacts

The wide variety of operating conditions results in a wide variety of materials used to meet the design needs of the engine Aluminum and titanium alloys, plastics, and resin-matrix graphite composites are frequently used in the fan and the engine nacelle The HPC uses titanium alloys, nickel-base superalloys, such as Inconel 718, and steels, such as M152, 17-4PH, and A286 The combustor requires heat-resistant nickel or cobalt alloys, such as Hastelloy X or Haynes 188, and stainless steels for fuel tubing The turbine sections rely on cobalt and nickel superalloys, such as Inconel X750, MAR-M-

509, René 77, René 80, René 125, and advanced directionally solidified and single-crystal alloys Often the design demands on materials for jet engine components will not permit substitution of materials for wear purposes, so a number

of surface coatings and treatments are employed for wear protection

The different operating environments and types of materials in each section of the engine result in a variety of wear types, including fretting, impact, adhesive, high-speed and oscillatory sliding, oxidational, ingested particle erosion, and abrasive wear High-speed sliding wear occurs in rotating gas path seals Impact wear can occur in loose part assemblies

or blade midspan or tip shroud interlocks Fretting wear is frequently seen in blade dovetails Erosion occurs when dirt and sand particles are ingested with the air through the fan and compressor Bearings and gears can experience rolling contact fatigue High-temperature components can experience oxidational wear This article will discuss some of the most significant of these wear problems in relation to specific jet engine components

Gas Path Seals

A major area of wear in jet engines involves gas path sealing Such seals include blade tip seals, labyrinth seals, and leaf and spline seals Blade tip and labyrinth seal problems are concerned with clearances between rotating parts and their adjacent stators Engine efficiency is significantly affected by the amount of gas leakage over blade tips or through labyrinth seals In an ideal engine, the blade tips or labyrinth seal teeth would maintain minimum clearance with the adjacent stator surfaces at all points in the engine cycle In practice, the rotor parts and stator parts experience differential growth rates because of thermal gradients in the engine and much larger mechanical growth of the rotor than the stator because of centrifugal forces The results of these differences are depicted in Fig 2 When the rotor and stator diameters are plotted against time after the throttle is applied for takeoff, they are seen to experience a period of interference (pinch point in Fig 2) that causes wear The issue can be further complicated by the stator going out of round (Fig 3)

Fig 2 Rotor and stator growth rates as a function of time and engine throttle movements

Fig 3 Clearance change caused by rotor/stator eccentricity or maneuver deflections

Both design and material approaches are employed in combating wear of blade tips One design approach, known as active clearance control, applies heating or cooling to the stator to achieve a better match of the thermomechanical responses of the stator and rotor A second design approach for the out-of-round condition involves local arc grinding to remove casing material in the areas where minimum radii would occur A materials approach that has been used in recent years is called passive clearance control The casing is made of an alloy with a low coefficient of expansion, such as IN909, to achieve a more favorable overall thermal transient response match of the rotor and stator diameters The most commonly used materials approach involves the application of an abradable material to the stator The abradable material wears preferentially in a limited arc when the stator is out of round or when the rotor moves off center This results in local clearance increases during rotor/stator interferences instead of wearing the rotor and causing a 360° clearance increase An alternative materials approach is to apply an abrasive to the rotor, which machines the stator material, thus achieving the same result

The materials used for the abradable stator seals or abrasive rotor coatings vary by location in the engine Abradables can take several forms, including bonded elastomers, braze-attached sintered porous metals or honeycomb cells, or thermal spray coatings Some of the more commonly used abradables are listed in Table 1 These materials are designed to wear

in preference to the opposing blade tip or seal tooth They rely on low densities created by included porosity or friable structures with weak bonding between constituent materials Bill and Wisander (Ref 1) have provided a model for friable abradable seal materials In practice, however, wear usually occurs on both surfaces, necessitating periodic overhaul

Table 1 Commonly used abradable seal materials

Phenolic/carbon microballons Aluminum

80/20 nickel-graphite Porous Teflon Aluminum-silicon/polyester

Fan and booster seals

Ni-Cr-Al/bentonite Nickel-graphites (75/25, 80/20, and 85/15) Nickel-aluminum

Aluminum Aluminum bronze/nickel-graphite Ni-Cr-Al/nickel-graphite

Ni-Cr-Al/bentonite Hastelloy X open-faced honeycomb

High-pressure compressor seals

FiberMetal Co-Ni-Cr-Al-Y

High-pressure turbine seals

Bradelloy (Hastelloy X honeycomb + braze/nickel-aluminum

The abrasive materials approach has been used with success on rotating parts, allowing them to machine their own clearances and minimizing rotor wear The most commonly used abrasive is plasma spray aluminum oxide on seal teeth

or rotor lands Figure 4 shows a sector from an HPC rotor with two sets of seal teeth coated with plasma spray aluminum oxide The most common mating stator seal material for such applications is open honeycomb (Fig 5) Commonly used abrasive coatings for clearance control in jet engines include:

• Plasma spray aluminum oxide

• Entrapment-plated cubic boron nitride (Borazon)

• Entrapment-plated aluminum oxide

The abrasive coatings approach is usually combined with honeycomb or an abradable seal to improve the overall wear system for both surfaces

Fig 4 High-pressure compressor disk with seal teeth coated with plasma spray aluminum oxide between

stages Arrow indicates location of coated seal teeth

Fig 5 Open-faced honeycomb seal showing cutting by seal teeth

Blade Midspan Stiffeners and Tip Shrouds

Some fan, HPC, and LPT rotating airfoils (blades) require the use of either a midspan stiffener or a Z-notch tip shroud (often called interlocks) to prevent mechanical flutter of the aerodynamically loaded blades These must be designed so that the blades are sufficiently loose to allow easy assembly, but lock up into a solid stiffening ring as aerodynamic loads are imposed on the blades, causing them to untwist along the blade stacking axis These two requirements result in a combination of impact and sliding as the interlocking contact surfaces engage and rotate into position to form the solid stiffening ring

The impact loads imposed on the contact surfaces can be on the order of 7 to 70 MPa (1000 to 10,000 psi) and can cause severe wear damage to most materials suitable for use as blades Therefore, it is common practice to apply a wear material to the interlock contact surfaces These wear treatments are usually coatings on the order of 0.13 to 0.25 mm (0.005 to 0.010 in.) thick or welded hardfacing deposits up to 2.5 mm (0.100 in.) thick Much care must be taken in the design and assembly of alignment tolerances for interlocks to prevent excessive wear, chipping, or spallation of even the most successful wear treatments on the interlock contact surfaces

The materials used for fan and HPC blades with interlocks are usually titanium alloys, which have poor wear properties Most fan blade and HPC interlocks use thermal sprayed WC-Co coatings or brazed-on WC-Co powder metallurgy wear pads to prevent excessive wear The most widely used coating is Union Carbide's LW1N40, applied using a detonation gun (D-gun) Recent advances in thermal spray coatings have allowed the use of high-energy plasma spray WC-Co coatings, which hold promise for direct substitution, or high-velocity oxyfuel (HVOF) sprayed WC-Co coatings on titanium alloy interlocks The WC-Co coatings are successful in the titanium alloy interlock applications because of the high wear resistance of the tungsten carbide, adequate fracture toughness because of the cobalt matrix, high adherence on the titanium alloy substrates, and a good match in coefficient of thermal expansion with the titanium alloy substrate materials The typical range of temperatures for fan and HPC interlocks may vary from subzero to 95 °C (200 °F) in the fan and from about 40 to 260 °C (100 to 500 °F) in the HPC Fortunately, WC-Co coatings appear to retain sufficient low-temperature ductility and high-temperature oxidation resistance over these temperature ranges The formation of a wear glaze at the contact zones contributes to the good wear resistance of the WC-Co in these interlock applications

The LPT blade materials are typically nickel-base superalloys, such as René 77 or René 125, which usually possess fairly good sliding wear resistance However, they have inadequate wear resistance in the combined impact and sliding wear environment of LPT blade interlock contact surfaces Typical use temperatures for LPT interlocks are 540 to 925 °C (1000 to 1700 °F), so the oxidation properties of the alloys under the existing wear conditions also play a significant role

in their wear resistance Typical wear coating compositions applied by thermal spraying or weld buildup that are used for LPT blade interlocks include:

• Tribaloy 800 (plasma sprayed, welded, HVOF)

• Cost Metal 64 (welded)

• Chromium-carbide/nickel-chromium (plasma sprayed, HVOF, D-gun)

Most of these alloys are cobalt based for good wear resistance and benefit from the formation of cobalt oxide and/or spinel wear glaze films

Tribaloy 800 and Coast Metal 64 are the most commonly used LPT blade interlock coatings at GE Aircraft Engines Tribaloy 800, applied by thermal spraying or tungsten inert gas (TIG) welding, provides excellent wear resistance and oxidation resistance to about 840 °C (1550 °F) Above this temperature TIG-welded Coast Metal 64 provides better wear and oxidation resistance than Tribaloy 800 In general, wear coating performance for LPT blade interlocks correlates to the chromium content and use temperature, with better performance at elevated temperatures for coatings with higher chromium contents and better performance at lower temperatures for coatings with lower chromium contents

Mainshaft Bearings

The materials traditionally used for gas turbine mainshaft bearings are 52100 and M50 steels More recently, powder metallurgy (P/M) bearings and case-carburized M50NiL steel, a modified M50, have been introduced as race materials Several factors have contributed to this recent trend Newer gas turbine mainshafts operate at higher speeds This has pushed the bearing DN values (bore diameter in millimeters times shaft revolutions per minute) well past 2 million, which increases race hoop stresses and the hertzian contact stresses between rolling elements and the races

The higher hoop stresses can cause fracture of the 52100 or M50-type races because they are through-hardened materials, typically in the 50 to 60 HRC range, and thus have low fracture toughness At high DN values, this can become a fracture reliability problem for a statistically significant number of these bearings The higher hertzian stresses, approaching 2400 MPa (350,000 psi) for a 2.5 million DN mainshaft bearing outer race, also can cause significant reductions in rolling contact fatigue life This is undesirable, because changing a mainshaft bearing requires costly disassembly of the engine

Bearing races made from the new P/M alloys and forged low-carbon alloys with carburized surfaces do not have these shortcomings These materials are designed for high DN use and require special manufacturing processes GE Aircraft Engines has concentrated on a variation of M50 steel with reduced carbon and increased nickel to improve fracture toughness The race is then carburized to produce a fine dispersion of carbides for high hardness Compressive residual stresses are frozen in to the raceway surfaces to improve rolling contact fatigue, while the low hardness (<50 HRC) rare core material remains tough to deal with high hoop stresses

This M50NiL material with its finely dispersed carbides, as well as the fine-grained (to improve fracture toughness) P/M race materials, can suffer from a low tolerance to wear Wear can occur at the ball cage guide lands under marginal lubrication conditions even for "normal" bearing cleanliness operation The rolling-element cage shoulders are silver plated, which provides solid lubrication and low friction to prevent wear when direct metal-to-metal cage skidding occurs

on the cage guide land of the race This works quite effectively for 52100 and M50 steels However, the P/M and M50NiL steels sometimes experience rapid wear under similar operating parameters

Research by Budinski (Ref 2) has shown that the size, distribution, type, and volume fraction of carbides in tool steels can significantly alter their abrasive wear resistance, with coarser carbide grains having better resistance than finer carbide grains Thus, it has been suggested that the coarser carbide stringers in M50 or 52100 forged bearing races can adequately resist the initiation of abrasive wear, while the very fine; evenly dispersed carbides in M50NiL or P/M bearing races cannot

The abrasive particles found in the bearings originate in the oil supply system Sump castings, abrasively cut tubes, and grit-blast-cleaned parts in the bearing lubrication system all likely contain very fine alumina or silicon carbide contaminant particles Most of these are removed during cleaning prior to assembly or by in-line filtering, but some of the finer particle (<50 m) get through to the bearings even under the most stringent clean-room assembly conditions Once inside a bearing, contaminant particles can become embedded in the silver plating on the ball cage shoulders, where they protrude, causing abrasive wear to initiate during transient cage shoulder/race guide land contacts The abrasive particles soon become "capped" with race transfer material, and adhesive wear ensues

The high differential sliding speeds between the orbiting cage and race cause frictional heating, local oxidation, and carburization by oil coking of the thin metallic transfer layers Thus, a very hard abrasive transfer layer results and the wear process accelerates These deposits also increase friction Therefore, when cage-to-cage encounters occur, a more severe rebound force results from the skidding contact, generally increasing the cage orbiting and number of skid contacts and producing further wear

Combustor and Nozzle Assemblies

These engine components are subjected to a variety of severe wear environments Combustor hardware includes fuel nozzles, swirlers, and cowl damping wires, which experience relatively high temperatures (540 °C, or 1000 °F, and up) during operation Exhaust nozzle assemblies are characterized by many parts, such as pins, bushings, links, and overlapping flaps, which aid in motion of the nozzle to control engine thrust Some of these nozzle parts are directly in the hot gas stream and experience temperatures up to 815 °C (1500 °F); others are bathed in bypass cooling air and remain relatively cool (approximately 315 °C, or 600 °F) Combustor and nozzle assemblies experience large amounts of vibration from turbulent air flows both inside and outside the engine This vibrational/impact wear can cause significant material removal as well as high-cycle fatigue of some components The combined effects can cause liberation of hardware; in the case of the combustor, this will in turn cause severe damage to downstream components, such as turbine nozzles and blades

The majority of wear problems in both of these assemblies is cause by vibration and impact Because of the elevated temperatures in the combustor, oxidational wear occurs as scales are formed and subsequent chipped off by impact Contact pressures between parts are nominally low, but can be aggravated by high-frequency impacts, which may locally yield the materials Oscillatory sliding (galling) wear sometimes occurs on exhaust nozzle flaps as they are actuated during mission cycles over several hundred accumulated flight hours

The design of these components addresses temperature and fatigue concerns Both combustor and exhaust nozzle hardware are made from heat-resistant superalloy sheet materials, such as Hastelloy X, René 41, or Haynes 25 Because these materials vibrate in the turbulent hot gas stream, high-cycle fatigue life at elevated temperature is important The cooler sections of the exhaust nozzle sometimes use high-temperature titanium alloys, such as Ti-6Al-2Sn-4Zr-2Mo, to reduce engine weight and maintain mechanical properties at elevated temperature Many pins and bushings are manufactured from steel alloys, such as 17-4PH and A286 Coatings can be applied to problem areas on specific components, but they must withstand the application temperatures and not degrade the mechanical properties of the base alloy to unacceptable levels Therefore, specification of the material and/or coating can be a complicated process

In general, cobalt-base alloys, such as Haynes 25 and Haynes 188, tend to perform best at temperatures above 540 °C (1000 °F) in both sliding and impact wear Some nickel-base alloys, such as René 41, also possess good wear resistance at high temperatures Effective coatings for wear problems in these temperature regimes include Tribaloy 800 and chromium carbide/nickel chromium Tribaloy 800 derives its good elevated-temperature wear resistance from a hard Laves phase in a cobalt-base matrix This matrix produces a cobalt oxide, which provides lubricity to the interface Chromium carbide/nickel chromium derives its good performance chiefly from the hard carbide phase and the formation

of favorable oxide wear glazes Cooler titanium components in the exhaust nozzle generally have poor wear resistance and almost always require coatings for mating parts in relative motion Here, the coating of choice is generally WC-Co, which again derives its wear resistance from the hard carbide phase Oxidation of the carbide limits use of this coating to temperature regimes below 480 °C (900 °F) Because of the aggressive nature of the carbide, both mating surfaces should

be coated The steels used in the actuation systems for the nozzle flaps are usually ion nitrided to develop a hard case layer on the surface (hardness of up to 72 HRC can be achieved) For a particularly severe wear environment, ion nitriding may not provide sufficient wear protection, and chromium carbide or tungsten carbide coatings may be required

Dovetails