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400 Commonwealth Drive, Warrendale, PA 15096-0001 U.S.A. Tel: (724) 776-4841 Fax: (724) 776-5760

SAE TECHNICALPAPER SERIES 1999-01-0142

An Investigative Overview of Automotive

Disc Brake Noise

K. Brent Dunlap, Michael A. Riehle and Richard E. LonghouseDelphi Chassis Systems

Reprinted From: Brake Technology and ABS/TCS Systems(SP-1413)

International Congress and ExpositionDetroit, Michigan

March 1-4, 1999

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1999-01-0142

An Investigative Overview of Automotive Disc Brake Noise

K. Brent Dunlap, Michael A. Riehle and Richard E. LonghouseDelphi Chassis Systems

Copyright 1999 Society of Automotive Engineers, Inc.

ABSTRACT

Disc brake noise continues to be a major concernthroughout the automotive industry despite efforts toreduce its occurrence. As a major supplier of automotivebrake components, Delphi Chassis is continually investi-gating means to reduce disc brake noise. In this paper,experimental and analytical methods are discussed

which reduce the occurrence of automotive disc brakenoise. Three general categories of brake noise are dis-cussed. These categories are low frequency noise, lowfrequency squeal, and high frequency squeal. A generaldescription of all three categories and examples of rele-vant solutions are presented.

INTRODUCTION

The purpose for writing this paper is to give the reader apractical overview of automotive disc brake noise throughsolution technique examples. Over the years, disc brakenoise has been given various names that provide some

definitions of the sound emitted such as grind, grunt,moan, groan, squeak, squeal, and wire brush. In order tosimplify the discussion presented in this paper, brakenoise has been divided into three general categories.The three groups presented are low frequency noise, lowfrequency squeal, and high frequency squeal. Theexamples presented in this paper outline experimentaland analytical methods used to provide insight into noisereduction solutions.

LOW FREQUENCY NOISE

Low frequency disc brake noise typically occurs in thefrequency range between 100 and 1000 Hz. Typicalnoises that reside in this category are grunt, groan, grindand moan. This category of noise is caused by frictionmaterial excitation at the rotor and lining interface. Theenergy is transmitted as a vibrational response throughthe brake corner and couples with other chassis compo-nents.

The following section looks at techniques used to charac-terize this noise and to develop solutions. The specificnoise investigated is groan, but the techniques presented

can also be used as general guidelines for developingsolutions for other low frequency noises

FRONT DISC BRAKE GROAN The typical failuremode for front disc brake groan occurs at decelerationrates in the range of 5 to 20 ft/sec2. Lining temperaturerange from 150 to 250F. Vehicle speeds are between 10and 20 mph. Some brake burnishing is usually neces

sary before groan will occur. The noise is normally pro-duced through the entire stop in the most severe casesbut is generally produced during the middle to end of thebraking event. The phenomena of sustained front discbrake groan discussed here was produced using Non-Asbestos Organic (NAO) lining materials. It is not theintent of this paper to discuss why NAO lining materialsseem to be more susceptible to the groan phenomenabut rather to discuss the methods used to reduce its propensity for producing groan.

Test Procedure The most important aspect of performing a meaningful experiment, particularly when dealing

with brake noise, is to establish a repeatable test sched-ule. A 20 stop test incorporating the failure mode discussed previously was devised. Interior noisemeasurements were recorded using both an objectivemeasurement system and subjective means. Accelerometers were located on suspension and brake corner com-ponents. A snapshot of the groaning event wascaptured, and the acceleration levels were ranked todetermine areas where detailed operating deflectionshapes (ODS) would be generated. An ODS providesinsight into the location(s) where modifications to the sys-tem could be made which reduce its dynamic responselevels.

Test Results Interior sound pressure recorded on theAachen Head measurement system was analyzed usingthe Binaural Analysis System (BAS). Filtering techniqueswere used to determine the frequency content of the sustained groan noise event. The noise was determined tobe relatively broadband in nature between 100 and 500Hz. A specific noise peak was present at 240 Hz, buwas not a pure tone. Figure 1 shows the sound pressurespectrum averaged over the entire stopping eventAcceleration spectrum shown in Figure 2 correlates withthe sound pressure. The ODS performed on the fron

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corners is described in Figure 3. The motion of the brakecorner is rigid body twisting of the caliper housing, caliperbracket, and knuckle

Figure 1. Interior Sound Pressure Spectrum - SustainedGroan Event

Figure 2. Lateral Caliper Acceleration Spectrum -Sustained Groan Event

Figure 3. Operating Deflection Shape of Brake Corner-Rigid Body Lateral Twist

Solution Techniques Two solution paths exist foaddressing this problem. One consists of modificationsto the vehicle response and the other of reducing theforcing function. The response side of the issue wasaddressed by modifying the stiffness, mass and/or damp-ing of various components along the response path. Thiswas accomplished by changing materials, adding tunedmass dampers, and stiffening suspension componentsThis solution path produced no favorable results. The

modifications shifted the frequency of the noise as muchas 100 Hz, but the overall perception of the groan was thesame. Based on these results, the only practicaapproach was to reduce the forcing function. Theapproach taken was to develop a new lining material thatwould reduce the occurrence of groan.

Lining Study In order to identify the primary character-istic(s) of the lining material that causes groan, severalining materials were evaluated for groan propensity. Thepurpose of the lining evaluation study was to generateclues that would lead to potential lining properties and/oingredients that could be identified as the root cause o

groan. The most obvious property that historically hasbeen associated with noise is the lining output level. Figure 4 shows the results of groan evaluations on severadifferent lining compounds. The chart shows specifictorque for NAO lining materials and semi-metallic liningmaterials versus groan propensity. It can be concludedthat output is not a root cause because low output liningsexist which produce groan. This evaluation generatedclues that identified lining material ingredients that wereassociated with groaning and non-groaning linings.

Figure 4. Sustained Groan Lining Candidates

To better understand the interaction of the raw ingredi-ents on sustained groan, a Design of Experiment (DOE)was developed. The DOE was a four factor - two levedesigned experiment. The response variable used wasthe number of stops with caliper acceleration greater than1 g of acceleration. This response variable was determined to be representative of the level of caliper acceler-ations generated during customer complaint vehicleevaluations. The experiment was performed using twovehicles that established four independent data samples(two brake corners per vehicle). A twenty stop schedule

110dBSPL

rms

30

Mag (dB)

Hz500100 Hz

Power Spectrum 2

30dBg

rms

-50

Mag (dB)

Hz500100 Hz

Power Spectrum 1

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

1.5

NAO

A

NAO

B

NAO

C

NAO

D

NAO

E

Semi-MetF

Sem-MetG

NAO

H

NAO

I

NAO

J

NAO

K

NAO

L

NAO

M

Lining

SpecificTorque Vehicle 1 Left Corner

Vehicle 2 Left Corner

Vehicle 1 Right Corner

Vehicle 2 Right Corner

Groan

Yes

No

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was used to evaluated each vehicle. Figure 5 shows theresults of the full factorial experiments. The values tabu-lated are the average number of stops with caliper accel-eration level greater than 1 g (maximum value = 20). Itshould be noted that comments describing other typebrake noises are also presented. Consequently, a liningtype that had low groan propensity was not necessarilygood for other brake noises. Figure 6 summarizes theresults into a Pareto of Effects." Statistically significant

contrasts (i.e., ingredients or combination of ingredients)that influence the propensity for groan are abrasive, lube-abrasive-filler, and filler.

Figure 5. Sustained Groan Factorial Experiment Results

Figure 6. Pareto of Effects for Groan DOE(A=Abrasive,L=Lubricant,Fil=Filler,Fib=Fiber,C=Carbon)

LOW FREQUENCY SQUEAL

Low frequency squeal is generally classified as noisehaving a narrow frequency bandwidth in the frequencyrange above 1000 Hz yet below the first circumferential(longitudinal) mode of the rotor. Circumferential modeswill be described in the sections pertaining to high fre-quency squeal. The failure mode for this category ofsqueal can be associated with frictional excitation cou-pled with a phenomenon referred to as modal locking ofbrake corner components. Modal locking is the coupling

of two or more modes of various structures producingoptimum conditions for brake squeal.

CASE STUDY The case study presented was per-formed on a front disc brake system. The failure modefor low frequency squeal occurs during low brake decel-eration, 4 to 6 ft/sec2, vehicle speeds of 5 to 10 mph, andinitial lining temperatures of 30 to 40F. Squeal no longeroccurs after approximately 10 stops.

The purpose of this case study was to investigate solution techniques that address the dynamic responseissues and not the excitation issues.

Test Procedure The same procedure was adhered towith this type of noise as was discussed earlier with lowfrequency noise. A schedule was established thaincluded 10 forward and 10 reverse stops at speeds lessthan 5 mph with apply pressures of 100 psi. The stopswere performed after an overnight cold soak at tempera-tures of 30 to 40F. An objective measurement systemconsisting of four microphones located at each wheel anda fifth microphone located at the passenger front seat

location was used to identify the squeal frequencyAccelerometers were also mounted on both front calipersfor correlation purposes. In addition, structural dynamicmeasurements were taken on the caliper housing, calipebracket, knuckle, pads, and rotor to characterize damp-ing, stiffness, resonance frequencies and mode shapes.

Test Results Sound pressure data captured during asqueal event indicated that the squeal was in the fre-quency range of 2500 to 2600 Hz. Structural dynamicmeasurements taken on the rotor and caliper whileinstalled on the vehicle indicated that both had reso

nances in that frequency range. The mode shapes forotor and caliper are depicted in Figures 7 and 8, respectively. The rotor resonance occurs at 2640 Hz and is the3rd nodal diameter mode. The caliper mode occurs a2546 Hz and is a 1st bending about the bridge.

Figure 7. Rotor Mode Shape @ 2688 Hz - 3rd NodalDiameter

Fiber(+) Fiber(-)

Lube(+) Lube(-) Lube(+) Lube(-)

Filler(+)

Abrasive(+) 10.75 0 9.5 0

Abrasive(-) 12.75 19.25 1 9

Filler(-)

Abrasive(+) 0.25 0.25 0.5 0.25

Abrasive(-) 7.25 6.5 6.25 0.5

Squeal

Grunt

Squeal

Grunt

Number of stops with caliper acceleration > 1g ( Maximum =20)

0

5

10

15

20

25

30

A L*A*Fil Fil Fib L*A Fib*A C L*A*C Fil*C

Factors

FValue

* *

** = Statistically Significant at 95%

Confidence Level

+

+

-

-

+

-

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Figure 8. Caliper Housing Mode Shape @ 2538 Hz

Solution Techniques The solution technique developedfor this case study was to decouple the caliper and rotormodes. Due to design constraints and timing, rotor mod-ifications were deemed to be a more practical short termsolution, however, there is no reason that the calipercould not have been part of the solution. As with any

structural dynamic modification, the damping, mass, and/or stiffness of the rotor can be modified. The simplestmodification that could be made due to hardware avail-ability, was to substitute a material change for gray castiron. This material substitution was classified asdamped iron. The goal of this modification was obvi-ously to reduce the amplitude of the rotor response byincreasing its damping.

Rotor Study The damped iron rotor was tested onvehicles, and the low frequency squeal was eliminated.Further investigation into the mechanism that eliminatedthe noise indicated that damping did not play a role in the

solution. Frequency response measurements in Figure 9comparing gray iron versus damped iron rotors showthat the resonance at 2600 Hz was shifted down in fre-quency approximately 400 Hz. Damping values mea-sured at those particular frequencies are the same.Therefore, the modulus change of the rotor not its damp-ing characteristic was the key to the noise reduction.

Figure 9. Frequency Response Measurement of GrayIron versus Damped Iron Rotors

HIGH FREQUENCY SQUEAL

One of the most troublesome noise issues in the brakecommunity is high frequency brake squeal. In this paperhigh frequency brake squeal is defined as noise which isproduced by friction induced excitation imparted on cou-pled resonances (closely spaced modes) of the rotoitself as well as other brake corner components. It typically is classified as squeal occurring at frequencies

above 5 kHz. Several solutions have been developedover the years that can greatly reduce the propensity forsqueal, but no absolute solution has ever been com-prised. It is not the intention of this paper to be a vehiclefor proclaiming to have found the absolute solution, butrather to provide insight concerning the mechanism bywhich some solution techniques have been successful.

DISC BRAKE ROTOR STRUCTURAL DYNAMICS ANDHIGH FREQUENCY SQUEAL For a particular brakesystem there exists particular frequencies at which highfrequency brake squeal will commonly occur. These frequencies typically remain constant for a particular brake

rotor independent of the rest of the brake system. Thislends credence to the theory that the brake rotor is a controlling element in determining the squeal frequencyThere are many bending modes of the disc brake rotorthroughout the frequency range of squeal, but the squeafrequencies are typically coincident with the circumferential or longitudinal resonance frequencies. This will beexplained further with the following case study.

Case Study Sample brake squeal data acquired with adragging brake dynamometer is shown in Figure 10aThis figure displays the frequency and amplitude osqueal for each stop of a randomized pressure and tem-

perature braking schedule. From this figure it is seen thabrake squeal occurred at three distinct frequencies. Frequency response function measurements of the rotodisc in the tangential and normal directions are d isplayedin Figures 10b and 10c. Modal analysis performed uponthese and additional measurements identified the resonance frequencies and mode shapes of the rotor discNormal direction resonances of the rotor disc existhroughout the displayed frequency range starting with a2 nodal diameter mode near 1 kHz up through a 12 nodadiameter mode near 15 kHz. The modal density in thetangential direction is much lower. The first three circumferential (tangential) modes of the rotor disc consisting o1, 2, and 3 nodal diameters are identified by Xs in Figure10b. A circumferential mode of a disc can be viewed as acompression wave in the disc circumference similar to alongitudinal mode of a solid bar. Directions of motion fonormal and circumferential (tangential) direction modesof vibration are illustrated in Figure 11. The rotor disc frequency response function data illustrates the strong relationship between the resonance frequency of thecircumferential modes and the frequency of squeal. Inmost cases of high frequency squeal, the squeal frequency relates to a circumferential resonance frequencyof the rotor disc, not a normal direction resonance

-+ +

Freq Resp 2:1

1000g/lb

0.01

Mag (Log)

kHz50 Hz

Freq Resp 2:1 Production Gray Iron"Damped Iron"

Frequencies of Interest

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frequency of the rotor or a resonance frequency of theother brake system components.

(a)

(b)

(c)

Figure 10. Correlation of Brake Squeal Frequencies andRotor Resonance Frequencies a)Dynamometer Test Squeal Frequencies b)Tangential Direction Frequency Response ofRotor Disc c) Normal Direction FrequencyResponse of Rotor Disc

Figure 11. Illustration of the Direction of Motion forNormal and Circumferential Modes ofVibration

The frequencies at which high frequency noise occur are

related to the circumferential modes of the rotor disc, butthe presence of these circumferential modes do not initself insure that brake noise will occur. For noise tooccur, it is expected that cross coupling of a circumferen-

tial and normal direction mode of the rotor or a circumferential mode of the rotor and another mode of the brakesystem must occur.

DISC BRAKE ROTOR STIFFNESS, LINING STIFFNESSAND FRICTION A model was constructed to gain anunderstanding of the relationships between the stiffnessof the rotor disc, the stiffness of the brake pad, the coefficient of friction, and the resulting normal and tangentia

direction response levels. As shown in Figure 12, thebrake rotor disc was modeled in a moving referenceframe as a continuous beam with stiffness and dampingelements in the equivalent normal and tangential directions, and the brake pad was similarly modeled with stiffness and damping terms.

(a) (b)

Figure 12. Normal-Tangential Mode Stability Model a)Physical Interpretation b) Model Configuration

The model was exercised to generate a stability diagramof pad stiffness versus the coefficient of friction for constant values of rotor stiffness as shown in Figure 13Based upon this diagram, reductions in pad stiffness oincreases in rotor stiffness for the same coefficient of friction will move the operating point to a more stable regionSuperimposed on the diagram are current values of stiff

ness for a brake system with a known propensity for highfrequency squeal. The diagram suggests that decreasing the brake pad stiffness to rotor stiffness ratio of thisbrake system will reduce the squeal propensity.

Figure 13. Stability mapping of Coefficient of Frictionversus Brake Pad Stiffness for Varying Levelsof Brake Rotor Stiffness

DISC BRAKE ROTOR DYNAMIC STIFFNESS ANDSQUEAL As previously discussed, a parameter thatcan impact the propensity to squeal is rotor stiffness. In

Dynamometer Test Sound Pressure Level Peak Distribution

50

60

70

80

90

100

0 2000 4000 6000 8000 10000 12000 14000 16000

Frequency [Hz]

SoundPressure[dB]

1000g/lb

0.01

Mag (Log)

kHz160 Hz

Freq Resp 2:1 K_PY.DATTangential (Y) Direction Driving Point

8k

X X X

1000g/lb

0.01

Mag (Log)

kHz160 Hz

Freq Resp 2:1 K_PZ.DATNormal (Z) Direction Driving Point

8k

1 ND Circumferential Mode2 ND Normal Mode

F=k

PADROTOR

kRT

kRN

F=KRN(RN-RT)

Coefficient of Friction vs. Pad Stiffness

0

0.1

0.2

0.3

0.4

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0.7

0.8

0.9

1

1.0

0E+05

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0E+05

1.0

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5.0

0E+06

1.0

0E+07

3.0

0E+07

5.0

0E+07

7.0

0E+07

1.0

0E+08

Pad Stiffness

CoefficientofFriction

1.00E+06

1.00E+07

1.00E+08

Rotor Stiffness

UNSTABLE REGION

STABLE REGION

Current Configuration

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the general sense, increased rotor stiffness is direction-ally correct for reduction in squeal propensity. This canbe looked upon as increasing the mechanical impedanceof the rotor and therefore making it more resistive inresponding to input forces. An associated factor andpotentially more significant factor of increased rotor stiff-ness is the shifting of normal response modes and theresulting reduction in modal density. The reduction inmodal density can inhibit the cross coupling of the cir-

cumferential and normal modes of the rotor disc as illus-trated in Figure 14.

Figure 14. Sketch of the Effects of Decreased ModalDensity and Potential Coupling of Normal andTangential Rotor Modes

A case study was performed in which rotor discs of differ-ing dynamic stiffness were evaluated on a brake systemwith a known propensity for high frequency squeal.

Rotors of increased and decreased cheek thickness(rotor rubbing surface) were utilized to achieve rotor stiff-ness variations. The resulting sound pressure data frombrake dynamometer tests shown in Figure 15a revealedthe large influence that rotor dynamics can have uponhigh frequency squeal. Prior to the test, it was antici-pated the noise rankings would correlate to the overallstiffness levels and associated trend of modal density, butthe data showed a different result. Tests conducted withthe baseline rotor produced the most noise, and testsconducted with rotors of increased and decreased cheekthickness produced lower noise levels. Further analysisof these rotors revealed differences in dynamic stiffness

at the predominant squeal frequency. This dynamic stiff-ness variation is related to the modal alignment of normaland circumferential modes. The dynamic stiffness data isshown in Figure 15b. Though decreased modal density

of the rotor disc is directionally correct for brake squeareduction, in this instance modal alignment of the normaand circumferential resonance frequencies were moreimportant than modal density as a single parameter.

(a)

(b)

Figure 15. Rotor Dynamic Stiffness Study Resultsa) Dynamometer Sound Pressure Datab) Rotor Stiffness Quantification

DISC BRAKE PAD GEOMETRY The sensitivity obrake pad stiffness upon brake squeal has been dis-cussed, and a factor that can affect brake pad stiffness isbrake pad geometry. Brake pad geometry can also havean effect upon the brake pad pressure distribution. Toinvestigate potential relationships between brake squeaand brake pad geometry, a case study was performedwhere varied geometry configurations of the disc brakepad were evaluated. Testing was performing on a dragging brake dynamometer over a wide range of lining tem-peratures and brake apply pressures. The sketch inFigure 16a displays the various configurations of liningshapes that were evaluated, the rotation direction, andthe orientation of the applied moment to the caliper bythe braking function. The modified pad shapes wereequivalent to a 20% reduction in lining arc length from the

baseline configuration. The results of the tests are displayed in Figure 16b.

f

TangentialMode

Normal Mode

f, MODAL SEPARATION

Original

f2 > f

Tangential Mode

Normal Mode

f

Modified

dB

f

Dynamometer Front Brake Squeal Evaluation

0

10

20

30

40

50

Plus 33% Nominal Minus 33%

Rotor Cheek Thickness

SoundP

ressure@1

1kHz

abovebackground,dB

Dynamic Rotor Stiffness at 11 kHz Squeal Frequency

0

1

2

3

4

5

Plus 33% Nominal Minus 33%

Rotor Cheek Thickness

D

ynamicRotorStiffness,

N/mx1e8

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(a) Disc Brake pad Geometry Study Schematicof the Various Test Configurations

(b) Disc Brake Pad Geometry Study DynamometerTest Results

Figure 16.

Significant variation in the noise was observed for thevarious configurations. Configurations D & E providedsignificant noise reduction, and lining configuration Fexacerbated the noise. The variation in noise perfor-mance of configurations A, B, C, & D can be theorizedbased upon their potential to induce lining sprag on therotor disc. The tendency of each configuration to sprag isbased upon the inboard and outboard balance of thebrake pads, the orientation of the caliper moment, andthe resulting pressure distribution variation across the arclength of the brake lining. This is very evident in configu-ration C where the caliper moment will tend to increasethe brake lining pressure at the cut edges for the inboardand outboard linings. The increased noise level for con-

figuration F can not be as readily theorized. To explainthe results for this configuration and to further the under-standing of the other configurations, pressure measure-ments across the length of the disc brake lining arerequired.

DISC BRAKE PAD NOISE INSULATORS Brake noiseinsulators are commonly utilized in disc brake systems asa noise fix. The insulators are typically constructed oflayers of steel and viscoelastic material applied to theback of the disc brake shoe plate. Because of the layer-

ing of steel and viscoelastic materials, the noise insulatois often viewed as a constrained layer damping treat-ment.

There are typically two schools of thought regarding themechanism of brake squeal reduction with respect tonoise insulators. One is the theory of the gasket effecbetween the shoe plate and the caliper where the noiseinsulator affects the mechanical impedance between thebrake caliper and disc brake shoe. The other is the

damping viewpoint; whereby, the insulator is effectivebecause of the additional damping added to the brakeshoe and lining in its constrained layer format. The constructions of some brake noise insulators appear toemploy both methods. Each view has their proponentsand regardless of which is correct or more correct, theeffectiveness of the brake noise insulator is dependenupon pressure, temperature, and frequency. This dependence is a result of the brake noise insulator being con-structed of viscoelastic materials. The viscoelasticmaterials used in brake noise insulators are typicallyadhesives and rubberlike compounds. The dynamicmaterial properties of these types of materials can vary

widely with temperature and frequency. As a result theperformance of brake noise insulators, whether supposedly designed to mismatch mechanical impedance or toadd structural damping, can vary significantly in theexposed operating environment.

A sample of the variation of damping levels with tempera-ture is shown in Figure 17. The data displayed in this figure was acquired from a disc brake pad in a free-freestate with different noise insulators applied. As displayed, the level of damping in the assembly can changesignificantly versus temperature. This data suggests thanoise insulators should not be generically applied to

brake systems. If damping optimization is applied, thedamping in the brake shoe assembly should be optimizedfor the temperature range at which squeal occurs.

Figure 17. Variation of Composite Disc Brake Pad Shoe &Lining Damping Versus Temperature

An illustration of the potential benefits of a brake noiseinsulator upon brake squeal is shown in Figure 18. This

RotorRotation OB pad

IB pad

Grayedareaindicatesremovedliningmaterial.

LHS Caliper

Moment

Configurations:

Baseline

A: Trailing End Cut

B: Leading End Cut

C: Inboard LE &Outboard TE Cut

D: Inboard TE &Outboard LE Cut

E: TE & LE Cut

F: Center Slot

TE LE

Equal Pad Length Ex perimental Results

0

10

20

30

40

50

Baseline TE Cut LE Cut IB LE Cut;

OB TE Cut

IB TE Cut;

OB LE Cut

TE & LE Cut

(~chamfer)

Center Slot

SoundPressure

[dB]abovebackgroun

7 kHz

13 kHz

Composite Shoe & Lining Damping vs. Temperature3rd Bending Mode

0.000

0.010

0.020

0.030

0.040

0.050

0.060

50 100 150 200 250 300

Temperature [F]

LossFactor[eta]

With Insulator

No Insulator

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figure displays the sound pressure levels for a vehiclebuilt with a wide variety of brake noise insulators. Thisillustrates the potential gains or losses in noise benefitsthat can be achieved with a properly tuned brake noiseinsulator.

Figure 18. Brake Squeal Test Results with Various BrakeNoise Insulators

CONCLUSION

In this paper brake noises were classified into three pri-mary categories: low frequency noise, low frequencysqueal, and high frequency squeal. Low frequency brakenoise is typically between 100 and 1000 Hz and is mostoften related to the rotor disc and brake pad frictionalinterface. Low frequency squeal is categorized by noisesin the 1 to 5 kHz frequency range. These noises typicallyoccur at frequencies of coincident modes for brake sys-

tem components. High frequency squeal occurs in theabove the 5 kHz frequency range, and the frequencies of

squeal correlate to the circumferential resonance fre-quencies of the rotor disc. A review of each category ofbrake noise and its related phenomena are reviewedbelow.

Low Frequency Noise

1. Low frequency noise is caused by friction materialexcitation at the rotor and lining interface that istransmitted through the brake corner and coupleswith other chassis components.

2. Front brake groan was determined to be broadbandin nature between 100 and 500 Hz

3. Attempts to modify the vehicle response with variousstructural modifications were unsuccessful.

4. The appropriate solution technique for solving groanwas determined to be a lining material modification.

5. A lining study was developed that pin-pointed themajor factors in the lining composition whichreduced the propensity for groan. These factorswere additional filler, abrasive, and fiber, as well as areduction in lubricant.

Low Frequency Squeal

1. The failure mode for low frequency squeal can beassociated with frictional excitation coupled with aphenomenon referred to as modal locking of thebrake corner.

2. Modal locking is the coupling of two or more modesof various structures producing optimum conditionsfor brake squeal.

3. Low frequency squeal is classified as noise havingnarrow frequency bandwidth in the frequency rangeabove 1000 Hz yet below the first circumferentiamode of the rotor.

4. A case study on low frequency squeal noise was presented and solution techniques were developed. Thetechniques addressed the decoupling of the calipeand rotor modes.

5. The caliper and rotor modes were decoupled bychanging the rotor material from gray cast iron todamped iron. This material change shifted the rotoresonance frequency down by 400 Hz.

High Frequency Squeal1. The frequency of squeal coincides with the circumfer

ential resonance frequencies of the rotor disc.

2. High frequency squeal propensity can be related tothe rotor to pad stiffness ratio for fixed levels of fric-tion.

3. Increases in the brake rotor disc dynamic stiffness aproblem squeal frequencies can reduce squeal propensity.

4. Disc brake pad geometry and brake pad pressuredistribution can have an impact upon brake squeal.

5. Brake noise insulators have the capacity to providesignificant reductions in squeal level if properly optimized for squeal conditions.

ACKNOWLEDGMENTS

The authors would like to thank Robert Ballinger, TimothyGillespie, Robert Lamport, William Myers, Sanjiv Tewaniand Julie Biermann-Weaver, for their contributions to thispaper.

REFERENCES

1. Matsuzaki, Kikio and Izumihara, Toshitaka, Brake NoisCaused by Longitudinal Vibration of the Disc Rotor, Pape

No. 930804 SAE International Congress and Exposition

1993.

2. Nashif, Jones, & Henderson, Vibration Damping, Joh

Wiley & Sons, Inc., 1985.

3. Lewis, Thomas M. and Shah, Praful, Analysis and Contro

of Brake Noise, Paper No. 872240, SAE Truck and Bu

Exposition, 1987.

Front Brake Noise Level with Va rious Noise InsulatorsCity Traffic Squeal Schedule, SPL @ wheel, left-right avg .

0

10

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40

No

Insulator

Vendor

A

Vendor

B1

Vendor

B2

Vendor

B3

Vendor

C

Vendor

D1

Vendor

D2

SoundPressure[dB]abovebackgroun