64 2 Design Essentials—Friction Material Composite SystemBFMC Development Calls for the Following Considerations Friction coefﬁ-cient and its stability over a range of operating conditions. The mating part or thecontacting surface and the contact surface form a friction pair and the properties ofthe counterpart rotor in the case of a disc brake, inﬂuence the μ not only directly butalso indirectly because of complex interactions between material and counterpartsurface. The main material members are polymer matrix ﬁbrous ﬁller particulateﬁllers of metal or mineral and lubricants. The friction of a matrix and its stability upon the type of polymer used can beaffected by the degree of polymerization or cross linking induced. Oxidation duringchemical changes in the surface of the polymer is a result of service. Most commonly used polymers are straight or modiﬁed phenolic or other ther-mosetting resins and elastomeric polymers. Polymer blends are invariably used to bring up a compromise between thermalstability and modulus of elasticity. The advantages of a highly thermally stable poly-mer are offset by the disadvantage of its high modulus of elasticity. Among the resins epoxies with amides, and polyimides are a good possibility.It is up to the designer to bring up an advantage in price working on ratios whilemeeting the cost and meeting the technical requirements. The loss in friction ofthe less stable material is partly reversible for its μ increases when the operatingtemperatures are reduced. Fiber reinforcement has a primary role to maintain intrinsic friction coefﬁcientin a design, for instance in the case of asbestos, higher stable friction coefﬁcient ismaintained by virtue of its large area of contact. It deforms easily to give a largearea of contact and ﬁber bundles easily open up to give clean surfaces.In the Case of Other Fibers and Particulate Fillers Particulate ﬁllers: Generalrule is friction coefﬁcient of any ﬁller is approximately 10–15 % of its hardnesswhich is related to logarithm of their Vickers Hardness Number (VHN). Harder mineral ﬁllers: increase in friction of the material Alumina increases theaverage friction level but μ also increases considerably as the speed decreases dur-ing a single application. To avoid such side effects very careful modiﬁcation of othermembers is required. Softer metals usually increase μ but their effect decreases with the increase inoperating temperatures. The μ of metals vary from 0.30 in antimony to 1.00 foraluminium with MOS2 0.10 with graphite having a μ of 0.20. Adsorption of watervapor or degradation products on the surface of the graphite can cause friction tovary with temperature.2.1 Brake and Vehicle DataDesigning the friction material composite calls for the essentials integrated to thebrake and vehicle design data (Table 2.1). Friction material composite performancevaries from brake to brake and vehicle to vehicle. It relies more on what brake in
2.1 Brake and Vehicle Data 65operation in a given vehicle design model. Example duo servo mechanism, Hy-draulic mechanism, Air assisted hydraulic mechanism, vacuum brakes etc. will varythe performance of the friction material pad or a liner in different vehicle system.In order to design a good friction material composite for a given brake system ina given vehicle model the design inputs need validation and complete testing inthat respective vehicle model. It would be wise to understand and acquire the basicessential knowledge of the brake system and about the vehicle overall in order todesign the friction material composite system.Minimum Design Requirement for a Good Friction Material CompositeWhile we understand the braking needs and vehicle needs for a good frictionmaterial design, essentially the following computation needs to be understoodto calculate the kinetic energy absorption/work done and horse power calcula-tions/retardation force and ﬁnally the torque. In order to achieve the required fullydeveloped mean deceleration the relationship between coefﬁcient of friction μ fordifferent line pressures can be correlated with the torque. An interesting mechanismcontrolled by the material inputs indicates even a minimum μ of 0.32–0.34 at max-imum line pressure the material can generate adequate torque to give the requiredfully developed mean deceleration. Here the brake system plays a crucial role whenthe operating variables like cut in pressure, pedal effort at knee point deceleration,booster size all play a good role. Details of torque computation for a typical disc pad application are given forreference.2.1.1 Data Collection Before Attempting Any DesignTable 2.1 Brake and vehicle dataVehicle details Passenger car I Passenger car IIGross vehicle weight 1280 (kgs) 1443 (kgs)Maximum speed of vehicle (Kph) 200 200Classiﬁcation as per JASO C406 PA PARoll radius of tyre F : 0.279 m F : 86 % R: 0.279 m R: 14 %Brake front Ventilated disc Ventilated disc Type C caliper Type C caliperSize φ 256φ mm ∗ 24 mm thick 256φ mm ∗ 24 mm thickMean effective radius 95 mm 105 mmNominal Mu 0.42 0.42Pad area 35 cm2 /pad 45 cm2 /padWheel cylinder φ 51 mm 52 mmMaximum hydraulic pressure 85.0 kg/cm2 110 kg/cm2Pedal gain 4.10 4.50
66 2 Design Essentials—Friction Material Composite SystemTable 2.1 (Continued)Vehicle details Passenger car I Passenger car IIBooster ratio 3.15 5.00Overall gain 12.91 22.50Master cylinder φ 20.63 mm 22.22 mmCut in pressure 30 30Valve ratio 4.0 (0.25) 3.33 (0.30)Type PCRV PCRV Pressure controlled Pressure controlled Release valve Release valve2.1.2 Basic Engineering Calculations to Design Based on the Theoretical TorqueTable 2.2 Basic engineering calculationsVehicle model Passenger car I Passenger car II (1280) (1443)K.E. absorbed at 1 2 ∗ 9.81 ∗ (27.77)2 = 50310 Kgm 1 2 ∗ 9.81 ∗ (27.77)2 = 56717 Kgm100 Kph 100 kmph = 27.77 m/sec 100 kmph = 27.77 m/secArea/brake 2 ∗ 35 cm2 = 70 cm2 2 ∗ 45 cm2 = 90 cm2Kgm/cm2 of disc pads 309 kgm/cm2 270 kgm/cm2Horsepower 0.6g 0.6gcalculationAssume a ‘g’ 0.6g 0.6g(constant deceleration)Stop time from 100 27.77 0.6∗9.81 = 4.71 sec 27.77 0.6∗9.81 = 4.71 secKph (sec) or 27.77 mpsRate of work done 21633 4.71∗75 = 61.23 HP 24388 4.71∗75 = 69.03 HP(HP)HP/cm2 of disc brake 61.23/70 = 0.87 HP/cm2 69.03/90 = 0.76 HP/cm2for two pads (70 cm2 )μ value from dyno test 0.34 (assumed) 0.34 (assumed)from 100 KphTotal retarding force F = (W/g) ∗ a F = (W/g) ∗ afrom 100 Kph/0.6g 1280 9.81 × 27.7 9.81∗0.6 = 615 kg 1443 9.81 × 27.77 9.81∗0.6 = 694 kgRetarding force/Front disc brake at 615 × 0.86 × = 264.45 kg1 2(assuming a braking ratio of 0.86 in the front to 0.14 in the rear).Torque Kgm 264 × 0.279 = 73.65 kg (retardation force × rolling radius) With the above said information one can calculate the theoretical torque (Ta-ble 2.2) that the friction material design can generate. It can be tested and veriﬁedwith the dynamometer test.
2.2 Design Drawing as an Input from the Original Equipment Manufacturer 672.1.3 Limiting Brake TorqueLimiting Brake Torque computation is calculated for the cast iron disc material ofthe disc under GG classiﬁcation. For a given deceleration ‘g’ based on a given brak-ing ratio and inertia, limiting brake torque is computed as below and if it exceedsthe torque calculated based on ‘g’ and braking ratio then the disc size requires im-provement. Torque equation: (TE − TA ) · α · ABS M= 1 − exp − CP ·GBS × t 2 · π · n α·A a = 59.7 J/m2 s K Material constants for GG cP = 51 J/N K Material constant for GG cP = Speciﬁc heat storage capacity (J/N K) GBS = Weight of brake disc [N] TE = Final temperature [°C] TA = Start temperature [°C] a = Transmission coefﬁcient [J/m2 s K] ABS = Transmission surface [m2 ] t = Braking time [s] It requires extensive understanding of multiphase, phase transfer and mass trans-fer issues to decide based on the material classiﬁcation and for a rotor size based onthe above. It is dealt with in detail in Volume 2.2.2 Design Drawing as an Input from the Original Equipment ManufacturerThe brake design drawing with the pad/liner furnished by the OE manufacturer willbear the complete dimensions to scale, the brake system and the friction materialpad/liner in the case of automobile/brake block in the case of rail applications. Asample copy of the approved design drawing from the design department wouldgive the details of critical essential technical speciﬁcations for testing and data forapproval. Besides that, the dimensions, with the tolerance, is to be completed forseveral views of the component design. Generally with the launch of any new vehicle model the design validation proce-dure goes for the ﬁrst 1–2 years of ﬁeld performance. It will be subject to changesand validation again as the theoretical design evolved will bear some modiﬁcationsonce it comes into the ﬁeld with the other members of the vehicle working together. A good design should work well before the launch of the vehicle model andwould require only ﬁne tuning of its component members after it is launched. A typical design drawing is one approved by the design department of the vehiclemanufacturer that normally takes into account the braking and vehicle manufactur-ers requirements.
68 2 Design Essentials—Friction Material Composite System A typical Original Equipment design drawing for a friction material designshould bear the critical dimensions besides the overall dimensions to scale and spe-ciﬁc test requirements which are relevant to ﬁeld design requirements. Tool correction, changes are a part of prototype tool development, should ad-here to strict standard speciﬁcations and dimensions. It should be well within thetolerance limits.2.2.1 Brake and Vehicle Dataa) Vehicle manufacturerb) Gross vehicle weight in kilogramsc) Front axle weight/Rear axle weight in kilogramsd) Maximum speed of the vehicle (Kmph)e) Classiﬁcation as per/Tatas/AK Masters/SABS/JASO etc. f) Rolling radius of the tyre (mm)g) Braking ratio—Front to rearh) Front brake disc and caliper type like ventilated/solid caliper C type i) Size of the brake j) Piston dia. (mm)/No. of pistonsk) Mean effective radius l) Nominal μ j) Pad contact areak) Rear brake size l) Rear wheel cylinder diameter (mm)m) Maximum hydraulic pressuren) Pedal gaino) Booster ratiop) Overall gainq) Master cylinder diameter r) Cut in pressures) Valve ratio with type of valve With some of the above said inputs, braking ratio could be worked out as givenbelow: (values are assumed)Gross Vehicle Weight: 1345 KgsRolling radius: 0.262 m2.3 Braking Ratio 2 DR BERR BF R PR 2 × × × DF BERF BF F PF
2.4 Inertia 69 DR = Wheel cylinder dia.—Rear 17.46 mm DF = Piston diameter—Front 50 mm BERR = Mean effective radius—Rear 0.09 m BERF = Mean effective radius—Front 0.095 m BF R = Brake factor—Rear = 2.0 BF F = Brake factor—Front = 0.8 PR = PF up to 35 kg/cm2 and valve ratio is 0.4For PF of 77 Kg/cm2 PR = 35 + (77 − 35) × 0.4 = 51.8 Kg/cm2 Rear (17.46)2 2.0 0.09 51.8 = × × × Front 502 8.0 0.095 77 2842.4 = BR = 83.7 : 16.3 (F : R) 146302.4 Inertia W I= × (RR)2 × BR × 1/2 G W = Gross vehicle weight in Kgs G = Acceleration due to gravity (9.81 m/sec2 ) RR in meter BR: Braking ratio 1345 I= × (0.262)2 × 0.837 × 1/2 = 3.94 Kgm sec2 9.81 Disc Rpm (N ) 16.67 N= × V (in Kph) 2πRR = K1 × V V = in Kph RR = in meter K1 = 2×3.143×0.262 = 10.122 16.67Mean torque via stopping distance (SD) Work done or energy absorbed WD = 0.5 × I × (ω)2 WD = T × ΦTherefore 0.5 × I × (ω)2 =T ×Φ I = Moment of inertia ω = angular velocity Φ = Stopping distance in radians = S/RR S = Stopping distance in ‘m’ RR = Rolling radius in ‘m’ ×(ω)2 T = 0.5×IΦ
70 2 Design Essentials—Friction Material Composite SystemFriction coefﬁcient T =2×p×A×μ×r μ = (T /p) × K3where 1 K3 = 2 × r × A × Hydraulic efﬁciency T = Torque (Kg m) P = Pressure (Kg/cm2 ) A = Area of caliper piston (cm2 ) r = Mean effective radius of discNote: Hydraulic efﬁciency assumed as 100 % Constant keyed in for computation is based on for a given piston diameter.2.5 ConstantsVehicle speed to disc revolutions per minute (RPM) Disc (revolutions per minute) 16.67/2π × RR × V in Kmph K1 = V /2π × RR where RR is rolling radius Assuming 60 kph 16.67/(2 × 3.143 × 0.262) K1 × V K1 = 10.122 Hence disc rpm = 10.122 × 60 = 607 rpm K2 = Deceleration via torque T = I × angular deceleration Angular deceleration = Linear deceleration/roll radius Linear deceleration = (T × RR)/I Friction coefﬁcient μ (brake factor) = T /p × K3 K3 = 2/r (mean effective radius) × A (piston area) × hydraulic efﬁciency K3 = 0.501 ∗ T K4 = Disc drag/Normal load (brake input) Disc drag = Torque/Mean effective radius of disc K4 = Disc drag/Pressure × area of piston × 2 (input load − normal force)/ 0.501 = 1.05.2.6 Terrain/Landform Topography as a Design InputDifferent terrains with their topographical variations become a critical factor for adesign to be successful or a failure. Variations in terrain such as hills/valleys/plains/rugged terrain/hot/cold/moderate terrains all need to be factored while designing, asthey have serious implication on the frictional performance, high temperature wearand fade/recovery characteristics.
2.7 Contacting Surface—Rotor Disc and Drum Details as a Design Input 71 In the case of a hot desert with hot days and cool nights the outside temperaturevariations can severely harm the brake and similarly in a valley with continuoussnowfall for most part of the year and a hilly region with heavy rains throughoutthe year require very careful planning of the design of the friction material compos-ite. In my next volume details of the design for the terrain and climatic variationswill be dealt with in greater detail. Normally all terrain variations are factored forthe respective terrains in the test schedule while qualifying the brake. Additionallyvehicle testing and ﬁeld evaluation would give further leads in understanding thebrake systems if there are any speciﬁc issues to be addressed.2.7 Contacting Surface—Rotor Disc and Drum Details as a Design Input2.7.1 Friction Induced Changes at the Rotor SurfaceBrake discs are normally made from cast iron. Both a Vented disc and a solid discare widely used for commercial and technical reasons and each has its own uniquecharacteristics of performance and wear when we brake. The bulk microstructureof the rotor comprises of graphitic ﬂakes in a pearlitic matrix. Turning or grindingof the surface gives a surface ﬁnish. After such a ﬁnish the surface is grooved andshows a bright contrast. When friction material surface comes into contact withthe rotor while braking the rotor surface is covered with gray, sometimes brownishlayer when viewed through a microscope. The sites covering the friction materiallayer will not exhibit the grooves [39, 40].Typical Technical Speciﬁcations of a Rotor Whether It Is a Grey Cast Iron/orAlloys GG20 Cr Cu HC Carbon—3.70–3.90 % by weight Chrome—0.20–0.35 % by weight Copper—0.50–0.65 % by weight Brinnell Hardness HBS/750 = 205 ± 5Surface Treatment Given on the Rotor Surface Surface machining of the friction ring-ﬁne tuned. Roughness “R3z5” Zinc surface protection (Zn)—thickness 8 to 20 µm
72 2 Design Essentials—Friction Material Composite SystemWhen we study the microstructure of the rotor discs the cast iron substrate will yielda pronounced channeling contrast. Due to severe plastic deformation fragmentationdue to deformation is visible and is ﬁlled normally by the wear troughs. Friction materials for disc brake applications have to be designed to provide areliable friction behavior for a large variety of different stressing parameters such asvelocity, pressure, temperature and humidity. The designing of the friction materialportion are done in such a way that the desired properties are met. It is highly im-probable that the distribution of the various constituents can bring about a perfecthomogeneity in a millimeter or a micro scale. Generally the micro constituents tendto bind themselves with the macro constituents e.g. in the form of a coating on steelﬁbers and Sb2 S3 (antimonium trisulphide) or as a premix of MOS2 (molybdenumdisulphide), SnS (tin sulphide) and or Sb2 S3 with silicates such as biotite or vermi-culite. The macro constituents should be distributed evenly in the surface and thespacing between them will normally be several millimeters. The ﬁne microstructure and homogenous chemical composition of friction lay-ers on both pad and rotor suggest that the iron oxide contains inclusions of solidlubricants on a very ﬁne scale in the form of nanoparticles.2.8 Brake RoughnessDisc brake roughness, a rigid body vibration, is caused by brake torque variationsmostly at wheel rotation frequency. It is felt as a tactile pulsation by the driver whodrives, as it often feeds back through the brake pedal and also in the steering wheel.Both the driver and the passenger may feel brake roughness through vehicle vibra-tions. It also occasionally causes sheet metal vibration .2.8.1 Roughness—Vibrational NoiseMany vibrations that are not due to brake roughness can occur at wheel rotationfrequency. For example, tire and wheel unbalance can cause tactile and visual vi-brations that may be sensed at the steering wheel. However, such vibrations do notrequire application of the brakes and tend to occur only at speciﬁc narrow speedranges (typically at 50–60 mph and 70–80 mph, sometimes as low as 30–35 mph).Poor suspension alignment, bent wheels, and some road surface irregularities canproduce vibrations that are similar to the ones caused by brake roughness. Some-times these may be more pronounced when the brakes are applied. Therefore, itis important to be careful in diagnosing and rating brake roughness on a vehicle.Proper vehicle instrumentation can be used to identify, quantify, and document brakeroughness test data. Disc brake roughness has been around for a long time, and has many root causes.Much has been understood on the causes,cures and on testing.
2.8 Brake Roughness 73 Roughness may show up only with cold brakes, sometimes with warmed brakes,or sometimes for all brake applications. Most vehicles have suspension and steeringsystems that get excited into greater vibration amplitudes at certain vehicle speeds(e.g., 30 mph). Prior brake usage history affects brake temperature distributions,their resultant brake thermal distortions, and thus also the tendency toward rough-ness. Experienced test drivers often choose a smooth road, then use speciﬁc vehiclespeeds and brake usage sequences to search for brake roughness. Different vehiclesuspensions, different steering systems, different caliper designs, different brake ro-tors, and different brake linings can all change the occurrence and severity of brakeroughness. New vehicle start-up time  is often a major concern about NVH problemsin general and brake roughness in particular. Prototype vehicles may have brake,suspension, and steering components that differ from the initial production parts. Attimes the new parts may appear to be better, closer to print nominal values, betterﬁnishes, etc. However, if they are different in any way, they may possibly showmore roughness. Even if the production parts are not changed, the higher number ofvehicles from full production may provide some with disc brake roughness. Some new vehicles may exhibit a brake roughness, including a pulsing feel onthe brake pedal, especially during a light brake application. These symptoms maydisappear after a few brake applications. If so, they probably resulted from con-tamination of the rotor surface. Local rusting of the rotor and/or oil/grease/paintcontamination of the rotor may be the causes. If the problem worsens with usage, a systematic diagnostic is required. Rotorsfrom problem vehicles should be measured for thickness variation (DTV), lateralrun out, and run out second harmonic. At a minimum, this should be measured atthe rotor mid-plane, but preferably also near the Outer Diameter and Inner Diame-ter. Vehicles vary in their sensitivity to rotor dimensional characteristics. Such sen-sitivity studies should be performed using production brake linings for the vehicle.Some brake linings have different elastic and frictional properties, so they inﬂuencethe rotor dimensional requirements for an acceptable brake rating. The brake linings used to evaluate brake roughness should be fully burnished. Toensure that the rating corresponds to steady-state customer usage conditions. Whenrating tests are run, the brake mechanic needs to be extremely careful to ensure thatneither the test linings nor the test rotor rubbing surfaces are contaminated by ﬁngercontact, or oil, grease, paint, or other extraneous materials.Brake Roughness—Mileage FactorSome semimetallic and Non-asbestos Organic brake pads cause brake roughness toworsen with time and vehicle mileage accumulation. This type of disc brake rough-ness results from a combination of abrasive pad surfaces and frequent highway/expressway driving. At least 2500 km of highway driving conditions, with a minimum of brake usage,is needed to develop high mileage roughness. Since it is mileage and usage sensi-tive, high mileage roughness may not appear until after 35000 km. Many roughness
74 2 Design Essentials—Friction Material Composite Systemsymptoms only show up after Fifteen thousand km on the highway. It is not uncom-mon for drivers to ﬁrst notice brake roughness after an extended driving vacation,since this type of driving hastens roughness occurrence. With higher mileage on thehigh way due to minimal usage of the brake roughness issue enhances. Abrasiveparticles at the brake pad surface can be the ﬁrst to contact the rotor. Under normalbrake pressures, and when the brakes are heated, most abrasive particles are em-bedded into the brake pad surface. This limits their abrasive action. However, whendriving at highway speeds with the brakes are released and cooled, a brake pad maygently and locally rub the rotor. The abrasive particles may ‘stand proud’ of the surface and dominate the contactsat such times. Eventually, this local contact of the rotor by the brake pad (especiallyby abrasive particles at the lining surface) will locally wear the rotor. This localrotor wear provides a rotor thickness variation, called RTV. RTV produces unevenbraking torques that may be especially noticeable on gentle brake applications. Theresultant periodic brake torque variations, and their associated brake pedal pulses,provide initial brake roughness.2.8.2 Rotor WearIt is always the localized rotor wear produces most brake roughness. This localwear almost always is produced during vehicle usage when the brake is released. Itis commonly worse when the brake pads are cool (below the binder resin glass tran-sition temperature). Under these conditions, a small amount of local brake draggingwears the rotor at the local contact site. With most disc brakes, this wear is con-ﬁned primarily to the inboard rotor face. The section Brake Design Factors providesa more complete explanation why the inboard brake pads cause most brake rotorRTV problems.2.8.3 Rotor Thickness Variation due to Excessive HeatOnce the rotor has developed signiﬁcant RTV, gentle brake applications provideuneven heating of the rotor. This becomes thermally induced RTV which increasesthe initial rotor RTV. Now the brake roughness is more severe. At higher speedswhen it reaches the point it excites suspension or steering component, the brakeroughness is observed to be higher.2.8.4 Disc Brake Roughness (DBR) MeasurementVehicle Roughness MeasurementsDrivers sense brake roughness through the brake pedal, steering wheel, seat assem-bly, ﬂoorboards, as well as through both visual and audible inputs. These are dif-
2.8 Brake Roughness 75Fig. 2.1 AFM picture of the roughness of the surface in a disc pad sampleﬁcult to quantify repeatedly. Most customer complaints on brake roughness comesfrom the drivers. From an experienced brake test driver roughness ratings are fairlyrepeatable, and are needed for ﬁnal vehicle ratings. Roughness of the surface as isseen in AFM pictures (Figs. 2.1 to 2.5).2.8.5 AFM—Brake Pad RoughnessCommon methods of vehicle roughness instrumentation are strain-gauged dragstruts and torque wheels. Both permit instrumented readings of brake torque av-erages and torque variations. All brakes have some torque variations, but not alltorque variations are at wheel frequency and large enough to be detected as brakeroughness. Instrumentation of the drag struts appears to offer both advantages anddisadvantages, compared with the torque wheels. Wheel torque are self-contained, not requiring application of the instrumentationdirectly to each test vehicle. However, they may provide a different wheel offset,mass, and stiffness than the OE vehicle has. They also may affect brake coolingrates and temperature distribution. This may affect the brake roughness amplitudeand occurrence conditions. When several test vehicles of the same make and modelare to be evaluated, wheel torques can be quite acceptable. If a number of samplesfor a particular vehicle is to be evaluated, for example to obtain an initial qualityrating, the use of wheel torque can be quite effective and efﬁcient. It should be
76 2 Design Essentials—Friction Material Composite SystemFig. 2.2 AFM picture of the roughness of the surface in a disc pad sample—friction materialportion of contactFig. 2.3 AFM picture of the roughness of the surface in a disc pad sample—friction materialportion of contact
2.8 Brake Roughness 77Fig. 2.4 AFM picture of the roughness of the surface in a disc pad sample—friction materialportion of contactFig. 2.5 AFM picture of the roughness of the surface in a disc pad sample—friction material ofcontact
78 2 Design Essentials—Friction Material Composite Systemremembered that torque variations do not necessarily correspond with the brakeforce output variations, such as seen by the drag strut, so torque data alone may notcorrelate well with driver ratings. Drag Strut Measurements: Strut instrumentation is particularly useful to char-acterize individual vehicles for roughness sensitivity. For example, a known set ofrotor/pad sets can be evaluated on a particular vehicle to establish that vehicle’ssuspension sensitivity to roughness. It is known that soft suspensions and soft strutbushings make vehicles more sensitive to brake roughness inputs. An instrumentedtest run with the same sets of brake rotors and linings on three vehicles each with adifferent suspension and/or strut bushing stiffness. The measured torque variationswere over three to four times greater on the vehicle with soft strut and suspen-sion bushings. It appears that strut bushing instrumentation is better for developingand tuning suspensions to minimize vehicles response to brake roughness inputs.Strut bushing test data (e.g., absolute amplitude or ratio of force variation to aver-age force) has provided a good correlation to experienced test drivers’ roughnessratings. Instrumented vehicle struts generally provide better vehicle roughness responsedata than instrumented wheel torque.2.8.6 Roughness Measurements in a DynamometerMost brake dynamometers have strain gauge torque sensors that can provide theneeded brake torque average and variation numbers. However, a brake dynamometerdoes not have the same brake mounting compliance as on a vehicle, and is connectedto the drive motor and load inertia by means of a drive shaft and couplings not as isin a wheel and tire. In its basic form, a brake dynamometer can provide useful dataon brake roughness. The ratio of peak-peak torque amplitude to average torque pro-vides a measure of the brake roughness input. A brake dynamometer can measuredifferences in this torque ratio for different test temperatures, different brake applypressures, and at different times during a simulated brake application. Brake rough-ness output, the observed vehicle response, varies substantially with this input. Bothbrake roughness input and output measurements are needed to determine the bestapproach to reduce brake roughness in the vehicle. Dynamometers normally do not provide information on how brake torque varia-tions may interact with such things as suspension geometry and component compli-ance (Fig. 2.6). Few brake dynamometers have the capability to include an entire ve-hicle corner-complete brake assembly, suspension, and structural components. Veryfew brake dynamometers absorb torque through tire/wheel assemblies. However, al-most any brake dynamometer can roughly simulate brake roughness deﬂections bythe addition of a spring element (even an actual strut bushing) to the brake tail stockreaction arm. This spring should be installed in series with the brake torque load cell. Thespring allows a test brake on a dynamometer to have nearly the same vibrational
2.8 Brake Roughness 79frequency as the wheel/tire/brake assembly on a vehicle. It is not known if dynowindup springs improve the correlation of roughness data from brake dynamome-ters to vehicle drivers. The important consideration is that the brake dynamometerreadily provides brake roughness input data, and can be modiﬁed to provide somesimulated output data. Vehicle roughness response characteristics, for the same brake input, may bequite different from one vehicle to another. It may be preferable to measure vehiclesuspension response versus frequency behavior using shakers at both front wheels tosimulate brake torque variations. This needs be done for each vehicle platform. Suchdata then can provide brake roughness torque variation bounds to achieve differentroughness ratings. Shows the drag strut response signal on three different vehicleplatforms with the same brake, under the same testing conditions. Most, but notall, of these differences were attributed to the strut bushing stiffness differences.Note the clear differences of signal amplitude and frequency that occurred beforethe brake was applied.DBR Disc Brake Roughness—CausesBrake roughness is excited by excessive brake torque variations. These may resultfrom one or more of several brake-related sources, most of which are ﬁrst order. By ﬁrst order, this means that a signiﬁcant event occurs only once per wheelrevolution. Examples are:1. Rotor thickness variations, RTV,2. First order brake Pad-rotor surface frictional variations,3. First order brake clamping force variations. Brake torque variations have their roots in brake design, materials, manufactur-ing, and usage history. However, there is more to brake roughness than simply theexcitation of the brake.Vehicle Design FactorsThe same brake hardware, installed in different vehicles, can provide large dif-ferences in reported brake roughness. Even when tested by the same drivers, theroughness ratings are clearly different for different suspensions and steering sys-tems. As with most vibrations, the brake roughness response is a function both ofthe brake excitation and of the vehicle system response to that excitation. Since thevehicle response to brake roughness inputs also is intimately tied to vehicle driveand steering behavior, brake engineers usually have to be content to address brakeroughness problems primarily through brake system modiﬁcations. Such constraintsmake roughness, ﬁxes difﬁcult to achieve on luxury vehicles with soft suspensions.This report does not address vehicle suspension and steering design changes to re-duce observable brake roughness. However, the non-brake contributions to reportedbrake roughness problems should be recognized.
80 2 Design Essentials—Friction Material Composite System2.8.7 Brake Design Factors—Sliding CalipersMost disc brakes today have sliding calipers, either pin or rail slider types, withpistons that use their seals for retraction. With such a design, if the outboard brakelining starts to drag against the rotor, its caliper readily moves over to reduce thecontact travel to a minimal value. This happens because the stiffness of the caliperassembly is high and its sliding force is low. On the other hand, if the inboard brakelining drags against the rotor, the caliper piston (suspended by the rubber seal) has alow stiffness, so it can move readily. The predisposes sliding caliper brakes towarddragging of their inboard linings. This is further biased by the normal displacementof the rotor, during cooling, toward the inboard lining. The bottom line is that slidingcaliper disc brake designs have an inherent tendency toward dragging of the inboardlining. The caliper piston travel, using its seal for a spring, may be 0.0020 to 0.003 fora dragging brake with rotor runout. The rotor contact, as might be expected, is alongthe runout ramp before, and after the point of maximum runout. Measurements tendto show brake dragging contact from about 60 to 80 degrees before maximum runoutto about 10 to 60 degrees after maximum runout on the inboard rotor face. When therotor has a high runout, the worn zone usually stays within 0.0015 of the maximum.This makes the worn zone narrower, with a resultant sharper brake torque pulse fromthe RTV. Brake lining drag wear is typically only on the inboard side of the rotor for mostpin and rail slider caliper designs. When outboard wear is found, it normally is onlya fraction of that found on the inboard side, and 180 degrees offset in location. Roadcrowns tend to provide a greater contamination to the right side brake assembly invehicles with right hand trafﬁc. Typically we one would ﬁnd more contamination-based rotor wear to show up on the right side rotors.Fixed CalipersFixed calipers can and do get RTV wear on both sides of the rotor. If the brakelinings are abrasive, the outboard wear can be about the same as that of the inboard.Fixed calipers are less common than sliding calipers. They tend to be used withrotors that have less runout, less tendency toward distortion, and are likely to havesuspension systems that are insensitive to brake roughness. There is not a great dealof data available on ﬁxed caliper disc brake roughness. Old data, from early Lincolnand Thunderbird ﬁxed caliper disc brakes, indicated their roughness was more notedwhen very hard brake lining were used, and when the vehicles were driven in regionswhere abrasive road dust was prevalent.2.8.8 Thickness Variation due to Manufacturing ReasonsSince disc brake roughness is directly related to brake torque variations, it is logicalthat variations in the thickness of a rotor, called thickness variation or RTV, would
2.8 Brake Roughness 81be important. Most caliper disc brakes have a very limited tolerance for RTV beforethe brake roughness becomes unacceptable. For this reason, disc brake rotors aregenerally machined on both rubbing faces at the same time. This may involve astraddle cutting on a lathe, or a grinding operation that machines both faces on thesame setup. Since cutting tools and grinders have some compliance, it is importantthat the roughing operations provide a minimal level of RTV as well.RunoutDisc brake rotors have some runouts without exception. It is not possible to elimi-nate all runout, since this involves bearing machining, bearing seats, bearings, ma-chine setups, and so forth. A small amount of runout generally will not induce adetectable brake roughness, at least initially. Large amounts of disc runout requirethe caliper and brake pad assemblies to move laterally with the runout, or the brakeclamping forces will vary with angular position. If it does, the brake may developroughness immediately due to the brake force variations. It also may develop brakeroughness during a prolonged low-pedal-force brake application, for example dur-ing a slowing for a freeway exit or a downgrade. During such braking, the rotorwill become heated unevenly as a result of the uneven clamping forces. This unevenheating of the rotor can increase the rotor runout and provide a signiﬁcant increaseof rotor thickness variation as well.Mass ImbalanceRotors may have castings that provide uneven mass distributions with angular posi-tion. These will respond to an even heating from brake application with an unevenchange of thickness. This thermally induced TV tends to be self perpetuating, onceinitiated. Ideally, the rotor contact faces should not vary in thickness with angular position.Residual StressSome RTV change after machining is possible if the gray iron casting is not stable.Initial heating of the rotor has been reported to produce permanent changes of mostdimensions, with runout changes being larger than those for RTV.Surface TextureUneven or irregular surface texture is not often a source for disc brake roughness.However, the initial roughness rating for new vehicles has been found to be sensi-tive to grinder alignment and bearing effects, when they produce an uneven surface
82 2 Design Essentials—Friction Material Composite Systemtexture on the rotor. Lathe turning is not known to produce uneven surface texture,but poor casting, with porosity, hardness, inclusion, free ferrite variations can resultin ﬁnished rotor rubbing surfaces that vary with angular position.Coatings on the SurfacesRotors at times are given a surface treatment, for example to provide rust protec-tion. It is important to be aware that any coating that affects the friction level, or isdepleted through wear, be thoroughly tested for its effect on brake roughness andvalidated. While such coatings may only temporarily affect brake roughness, anydetectable adverse effect can elicit a strong negative ﬁrst impression by the cus-tomer.Reasons—Usage RelatedBrake Parasitic Drag Wear When a disc brake is released, the piston seal roll-back retracts the piston several thousandths of an inch. This small retraction isneeded to minimize brake pedal travel for initial lining contact. However, the smallseal roll-back may result in some local brake pad contact when the brake is released.This is called parasitic drag. Normally this drag is small, about the same as wheelbearing or seal drag. However, it can have serious brake roughness consequencesunder certain circumstances.2.8.9 Abrasive Brake PadsSome brake linings contain abrasives as a part of their composition. For example,many semi-met friction materials contain fused magnesium oxide of a particle sizethat can be abrasive to the rotor. Abrasive materials also may occur as unwanted,‘tramp’ constituents in brake linings/pads. Silicon carbide is a well known abrasivematerial that may be found in synthetic graphite. Accumulated surface materials,such as road dust or rotor rust particulate may collect on the brake lining rubbingsurfaces. Some abrasive material is possible in and on a brake lining surface. Theharder the brake lining matrix, the smaller the abrasive particles need be to wear therotor. For this reason ‘soft’ brake linings and warm brake pad tend to wear the rotorsmuch less aggressively. When the brake is nominally released, but with some parasitic drag, the brakepad surface periodically contacts a portion of the rotor surface. If the brake padsurface that contacts the rotor is abrasive, even this light contact may result in alocal rotor surface wear. Such wear results in usage TV. This wear is generally onthe inboard face of the rotor. The reason for this is that most disc brake calipers
2.8 Brake Roughness 83have their pistons on the inboard side, with a sliding mechanism of the caliper forthe outboard shoe loading. If abrasive contamination comes from road dust, the outboard rotor wear canbe much less, especially for vehicles with closed disc wheels or with closed wheelcovers, These minimize abrasive particulate entry to the brake. If spoked wheelswith large opening are used, both rotor faces may wear about the same from road-borne abrasive contamination (see Fig. 2.7), which will lead to disc scoring. Roadcrowns tend to provide a greater contamination to the right side brake assemblywhere there is right hand driving. Consequently, the typical situation is for morecontamination-based rotor wear to show up on the right side rotors.Runout Induced RTVRTV change of 0.0015 as measured are common on a large passenger car rotor dur-ing a slow brake application from 100 to 60 kmph. The initial thickness variationnormally could be under 0.0001 on the rotor, such an instance cannot be attributedto braking roughness. It is important to remember that a caliper disc brake is alwaysunstable in terms of thickness variation. The ﬁrst and second order components ofthe rotor runout result in some variation of brake lining contact pressure when somethermally-induced thickness variation starts. Any RTV will tend to increase withtime during a prolonged light brake application. Some brake linings are more likelyto generate regional hot spots, and associated brake roughness. Soft (in compres-sion) brake linings are better than rigid materials, as they tolerate the runout withless frictional force variation. Lighter weight calipers and free moving calipers sim-ilarly reduce the vehicle sensitivity. On some of the brakes, an increase of hydraulicbrake line size also reduce the runout induced TV. It could lead to softening of thecaliper piston with the larger hydraulic line, reducing the brake lining drag forcevariation with runout.Understanding of brake roughness In the case of steering wheel response,thickness variation phasing controls the magnitude of the steering wheel response.From highway to light steady braking steering wheel oscillation becomes worse dueto roughness. Fixing one side will stop all steering wheel oscillation, but brake pedal pulsationwill continue. Normally Drivers complaint when both rotors have excessive thick-ness variation as it causes steering input. This phasing effect caused the roughnessto vary substantially, even if the test conditions are repeated. Composite stamped ro-tors give poor runout than the cast rotor. With cold brakes especially in highway typeusage will result in increased roughness. Abrasive content will cause increased facewear which is seen in semimetallic formulations. Highway usage-induced rough-ness does not occur when drivers use brakes enough to keep the linings, pads abovetheir glass transition temperature of 84 °C. Varying suspensions and strut bushingswill vary the disc brake behavior.
84 2 Design Essentials—Friction Material Composite SystemFig. 2.6 Vented rotor discwith caliper mounting with anadaptor in a typicaldynamometer test setupFig. 2.7 Scored vented discafter undergoing severalcycles of thermal history The drag strut bushing spring rate was the greatest single variable that affectsroughness. Initial rotor thickness variation will be over seven times greater in ef-fect than initial rotor runout. However, runout was the greatest root source for highmileage thickness variation increase and high mileage driver complaints of rough-ness. Uneven coating on a new vehicle wherein the rotor roughness is seen until itis wiped off during repeated braking. Issues of disc rotor in contact with the friction material surface are more relatedto compressibility of the product mix formulation in question.
2.8 Brake Roughness 852.8.10 Metallographic Studies on Grey Cast Iron Samples of the DrumMetallographic studies such as macroscopic examination, microstructural analysisand hardness testing could reveal if there is any abnormality in the drum liner contactwhich is very critical as a mating part. Both scored and unscored drum samplescould be taken for studies of metallography. Macroscopic examination of the grey cast iron with and without scores couldbe observed visually. Both the surfaces with and without scoring samples could bemetallographically polished. Micrographs could be taken on samples with etchedand unetched conditions . At different points over the surface it could be stud-ied. Some locations may reveal graphitic ﬂakes of type E having interdendritic segregation with preferred orientation in a scored drum. Repetition of other loca-tions of similar patterns are sometimes seen. In an unscored sample locations mayshow tendency for growth of graphite ﬂakes of a particular type. Normally size ofthe graphite ﬂakes will correspond to ASTM designation A 247 Plate I. Etched scored sample will show pearlitic with a few grains of ferrite and some-times steadite could be noticed. In an unscored sample microstructure reveal a re-solved pearlite matrix and white etching steadite. Different locations can reveal sizevariation of steadite grains. Vickers hardness testing with 5 kg load may reveal some informations on thelocations where metallographic studies are carried out. There will be variation ofhardness on both scored and unscored samples tested at different locations. Graphite ﬂake size and distribution depends more on cooling rate and the thick-ness of the casting. The chemical composition of the cast iron also will inﬂuencethe nature of graphite. Sometimes across the section the cooling rates will not beuniform . Higher content of steadite indicates higher phosphorous content inthe sample.