M. Triches Jr Et Al Reduction Of Squeal Noise From Disc .

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M. Triches Jr et alM. Triches Jr, S. N. Y. Gergesand R. JordanFederal University of Santa CatarinaMechanical Engineering DepartmentCampus Universitario, TrindadeP.O. 47688040-900 Florianópolis, SC. .ufsc.brReduction of Squeal Noise from DiscBrake Systems Using ConstrainedLayer DampingSqueal noise generation during braking is a complicated dynamic problem whichautomobile manufacturers have confronted for decades. Customer complaints result insignificant yearly warranty costs. More importantly, customer dissatisfaction may result inrejection of certain brands of brake systems. In order to produce quality automobiles thatcan compete in today’s marketplace, the occurrence of disc brake squeal noise must bereduced. The addition of a constrained layer material to brake pads is commonly utilizedas a means of introducing additional damping to the brake system. Additional damping isone way to reduce vibration at resonance, and hence, squeal noise. The simulation ofbraking events in dynamometers has typically been the preferred insulator selectionprocess. However, this method is costly, time consuming and often does not provide aninsight into the mechanism of squeal noise generation. This work demonstrates the use ofmodal analysis techniques to select brake dampers for reducing braking squeal. Theproposed methodology reduces significantly the insulator selection time and allows anoptimized use of the brake dynamometer to validate selected insulators.Keywords: Brake, damping, squeal, noiseIntroductionDisc brake noise is an ongoing problem for the automotiveindustry. Brake noise is perceived by customers as both annoyingand an indication of a problem with the brake system. In most cases,this type of noise has little or no effect on the performance of thebrake system. However, its perception dramatically affects qualityand satisfaction ratings as well as warranty costs. This is the reasonwhy the automotive industry is looking for ways to control it.1Considerable effort has been directed at investigation andreduction of disc brake noise. Most of this work has been performedon problem brake systems whose design is finalized (Triches et al.,2002). In these cases, the only solution available is the applicationof noise control methods. As a consequence, add-on noise controltreatments have become a very common technique in reducing thebrake noise problem. However, the application of these treatments issometimes regarded as an iterative procedure, where the effects of ahuge matrix are evaluated on a structure experimentally.In most cases, the iterative procedure to select an appropriatenoise control treatment for brake noise problems involves the use ofan inertial brake dynamometer. This procedure, however, is costlyand time consuming, because of the interaction between theproperties of damping materials (i.e. loss factor and shear modulus)and the resonant response of the brake assembly (shoe and lining,rotor and caliper).In contrast, the design of effective noise control modifications toreduce the brake noise problem can be achieved efficiently usingexisting experimental techniques and methodologies. The first stepis to define the dynamic characteristics of the brake system in termsof noise generation, identifying the source and the mechanism of theaudible noise emissions (Papinniemi et al., 2002). Once thesecharacteristics are understood, a suitable damping material to reducea specific brake noise problem can be selected using experimentaltechniques and material damping knowledge.This paper is concerned with describing the application of modalanalysis tools and damping materials knowledge to select a suitablebrake noise insulator to reduce the squeal noise problem. Thismethodology is applied to a particular brake system and the resultsobtained are presented. This approach is validated through newdynamometer tests, with the selected damping material applied tothe brake system.There are several categories of brake noise that are classifiedaccording to the frequency of noise occurrence. Basically, there arethree general categories of brake noise: low frequency noise, lowfrequency squeal and high frequency squeal (Dunlap et al., 1999).Low frequency disc brake noise is a problem that typically occurs inthe frequency range between 100 and 1000 Hz. Typical examples ofnoise problems from this category are groan and moan noise. Thegeneration mechanism of this kind of problem is the frictionexcitation at the rotor and lining material, which provides energy tothe system. This energy is transmitted as a vibratory responsethrough the brake assembly and couples with components of thesuspension and chassis.Although the low frequency noise is an important problem forcertain types of brake systems, the most common and annoyingproblem is squeal noise (Dunlap et al, 1999). Squeal is defined as anoise whose frequency content is 1000 Hz or higher that occurswhen a system experiences very high amplitude mechanicalvibrations. There are two theories that try to explain how thisphenomenon occurs. The first one is called “stick-slip”. Accordingto this theory, squeal is a self-excited vibration of the brake systemcaused as a result of two factors: the static friction coefficient isgreater than the sliding friction coefficient; the relationship betweensliding friction coefficient f and relative sliding velocity Vr isδf δVr 0 . However, this theory cannot explain why thetendency of squeal is different when the same friction couple pair(rotor and pads) is used in different brake systems (Chung et al.,2001). Therefore, a second theory, called “sprag-slip”, wasdeveloped. It demonstrates that the self-excited vibration of thebrake system and the high levels of vibration result from animproper selection of geometric parameters of the brake system. Inthis case, two system modes that are geometrically matched movecloser in frequency as the friction coefficient increases. These twomodes eventually couple at the same frequency and matching modeshapes, becoming unstable (Dihua and Dongying, 1998). Boththeories attribute the brake system vibration and consequent noise tovariable friction forces at the pad-rotor interface. These variablefriction forces introduce energy into the system. During the squealevent, the system is not able to dissipate part of this energy and theresult is the high level in the amplitude of vibration.Paper accepted August, 2004. Technical Editor: Atila P. Silva Freire.340 / Vol. XXVI, No. 3, July-September 2004ABCM

Reduction of Squeal Noise from Disc Brake Systems Using These two theories have been investigated and discussed byresearchers, but previous brake squeal experience and the majorityof research literature considers the geometric instability to be themajor mechanism of generation of brake squeal (Abdelhamid et al.,2001).There are two types of brake squeal: low frequency and highfrequency squeal. The difference between them is the mode shapesinvolved in the modal coupling mechanism. For the low frequencysqueal, the modal coupling occurs between the out-of-plane modesof the rotor and bending modes of the brake pad. For the highfrequency squeal, the modal coupling occurs between the in-planemodes of the rotor. The brake rotor is much stiffer in the in-planedirection than in the out-of-plane direction. Therefore, the resonancefrequencies of the in-plane modes of the brake rotor are higher thanthose of the out-of-plane (bending) modes. Figure 1 shows thecoupling possibilities between brake components.Figure 2. Components of an inertial dynamometer.Measurements were taken for brake temperatures between 50and 300 ºC. The braking pressure varied from 5 to 40 bar. Theacquisition system stored the data of the brake events in blocks withthe same temperature and pressure conditions, allowing acomparison between the different conditions to find regions oftemperature and pressure where the noise occurs.Each brake event lasts approximately 10 seconds. During thisperiod, data were sampled in a certain number of autospectra of themicrophone signal. The Sound Pressure Level (SPL) reported foreach brake event is the maximum value measured among allautospectra.Figure 1. Coupling possibilities between brake components.Typically, the high frequency squeal occurs for frequencyranges between 8 and 16 kHz, while low frequency squeal occursbetween 1 and 7 kHz. Since the human ear is most sensitive to thefrequency region between 1 and 4 kHz, the low frequency squeal isconsidered the most annoying type of brake noise.For these reasons, this paper attempts to evaluate controlmethods for the low frequency squeal problem using dampingmaterials, and also procedures for selecting an appropriate dampingmaterial to fix a particular problem of low frequency squeal in anexisting disc brake system.Characterization of Brake Noise GenerationPerhaps one of the most important pieces of information frombrake systems is the characterization of the noise generation viadynamometer or vehicle testing. Tests on vehicles are veryimprecise, because it is impossible to control the variables likevelocity, brake pressure and temperature in order to produce resultsthat represent the characteristics of brake noise generation. On theother hand, dynamometer tests allow us to approach the realbehavior of a brake system in practice, controlling parameters suchas rotation, braking pressure and temperature while recording thesound pressure level and frequency of noise occurrences. The discrotation is achieved by electric motors connected to a shaft withinertial wheels, simulating the inertia effects of the vehicle. Figure 2shows of the components of an inertial brake dynamometer.J. of the Braz. Soc. of Mech. Sci. & Eng.Figure 3. Noise occurrences obtained with dynamometer test.Figure 3 shows the results obtained for the baseline brakesystem, i.e, without any modification to its components. Thedynamometer results show a strong noise frequency around 7 kHz,together with other peaks across the whole frequency range.However, the number of occurrences and the amplitude of the SPLpeak indicate that the noise at 7 kHz is the most important problemin this particular brake system under investigation. Figure 4 shows anoise map obtained with the dynamometer test. It can be seen thatthe highest SPL peak occurs for a frequency around 7 kHz, for atemperature around 150 ºC and for a pressure of 25 bar. This kind ofCopyright 2004 by ABCMJuly-September 2004, Vol. XXVI, No. 3 / 341

M. Triches Jr et alinformation is important to identify which class of noise problem isresponsible for the high noise levels from the brake system.Figure 5. Example of excitation autospectra used in the modal testingprocedure.Figure 4. Noise map for baseline brake system.The next step is to determine the modal behavior of each brakecomponent to verify whether any component has a resonancefrequency near 7 kHz, and more importantly, to detect potentialmodal couplings between them. Therefore, a modal analysisprocedure is applied for each component, i.e, rotor, pads and caliper,and the natural frequencies and mode shapes of these componentsare obtained.Characterization of the Dynamics of Brake ComponentsSqueal noise occurs only when the brake system componentsdemonstrate resonance vibrations (Boss and Balvedi, 2001).Therefore, it is very important to determine the modal behavior ofthe components to understand the problem.Modal analysis of individual components allows us to gain aninsight into potential coupling modes, which is, as mentionedbefore, one of the causes of squeal noise generation. In order toobtain the modal parameters of brake components, each one ismodeled through a mathematical mesh to represent its geometry.Pad, rotor and caliper Frequency Response Functions (FRF s)were measured by exciting each component with an impact hammerand measuring the acceleration response with a small accelerometer.Then, the FRF s were processed by CADA-X software in order toidentify the modal parameters, i.e, resonance frequencies, modalshapes and damping values.Modal Analysis of Brake PadsThe excitation was provided by a small impact hammer (PCB086D02), with sensitivity of 22 mV/N and with a hard tip to achievefrequencies up to 16 kHz (see Figure 5). In order to avoid errors dueto the effect of transducers in the dynamic properties of the brakepads (additional mass), a light small accelerometer (PCB 352B10),with sensitivity of 10 mV/g, was used to obtain the accelerationresponse. The accelerometer was kept at a fixed point (see Figure 6)and the excitation was applied at all points (roving method). Theanalysis was carried with 4096 data points, 16384 kHz of maximumfrequency and frequency resolution of 4 Hz. An exponential timeweighting function (window) was used for the response signal (5%decay and damping correction) and a transient window was used forthe force signal. The modal parameters were extracted using thetime domain method (least squares complex). Figure 7 shows themode shapes and Table 1 presents the resonance frequencies and thedamping loss factors obtained for the brake pad.Figure 6. Mesh used for modal testing procedure of the brake pad.The modal analysis of brake pads is perhaps the most importantprocess to understand in order to find solutions for the disc brakenoise problem. Some properties like loss factor, natural frequenciesand mode shapes of brake pads are crucial in defining which type ofbrake noise problem may occur.The brake pad was supported by two slender cables in order tosimulate a free-free boundary condition. The free-free conditionallows the structure to vibrate without interference from other parts,making the visualization easier of mode shapes associated with eachnatural frequency. Moreover, in this case, the rigid body oscillationfrequency of the assembly (suspended pads) is much lower than thefirst natural frequency of the structure (pad). For instance, theassembly rigid body frequency is around 5 Hz, while the firstnatural frequency of the brake pad occurs around 2600 Hz.Figure 7. Mode shapes for brake pad.The mode shapes for the brake pad are very similar to bendingand twisting modes of beams. The pad length is longer than thewidth. As a consequence, the bending modes along the longer edgeoccur first. From the modal coupling point of view, the bendingmodes are more important than the twisting modes. In most cases,342 / Vol. XXVI, No. 3, July-September 2004ABCM

Reduction of Squeal Noise from Disc Brake Systems Using the modal coupling occurs between pad and disc bending modes,since the disc does not have a defined shape for twisting modes.Table 1. Modal parameters obtained for brake pad.VibrationMode12345ResonanceFrequency (Hz)26203757665071538623Mode shape1 st bending2nd bending3rd bending1 st twisting2nd twistingDamping LossFactor (%)0.6780.6460.6411.0520.448Modal Analysis of DiscIn the same way as that for the brake pad, the modal analysis ofthe disc was carried out. The mesh was constructed with 111 pointsto avoid space aliasing. During the measurements, the rotor wassupported by a foam block, in order to simulate a free-free boundarycondition. Experiments show that analysis with fixed boundaryconditions, i.e, disc fixed on the brake knuckle by bolts, generatesmode shapes very close in form to the mode shapes obtained for therotor in the free-free condition.The disc mesh represents only the border, because the disc hatand bolt area can be considered without influence on the couplingmechanism. Furthermore, the mode shapes found for these regionsare located at high frequencies, beyond the frequency range ofinterest for the analysis of potential coupling modes. Figure 8 showsthe modal shapes and Table 2 presents the natural frequencies anddamping loss factors obtained for the brake disc in normal direction.of this study, is sufficient. For this case, a modal analysis in normaldirection is enough.The pad presents less vibration modes than the disc, over thesame frequency range. Furthermore, the damping loss factor valuesassociated with the pad vibration modes are higher than those of therotor, because the friction material provides considerably moredamping than the cast steel used in the disc. As a consequence, thereis a tendency for the disc modes to be the major determinant of thesqueal frequency.Table 2. Modal parameters obtained for brake disc.VibrationMode1361013ResonanceFrequency (Hz)10902210360053207320ModeDamping LossshapeFactor (%)2nd bending0.2473rd bending0.1284th bending0.1085th bending0.1306th bending0.176As mentioned before, squeal noise usually occurs whenever anumber of brake components, such as pad and disc, start to vibratetogether, creating a coupled system mode. Considering the bendingmodes coupling, when the components have the same wavelengthand frequency, they will be geometrically matched and will vibratein phase (Fieldhouse, 1999). In this case, the friction damping isminimized and the system works as a loudspeaker, radiating sound.Analyzing the results obtained with the modal analysis of brakecomponents together with their geometry, it can be noticed that thethird bending mode of the brake pad and the sixth bending mode ofthe disc can couple and create a system mode. The third bendingmode of the pad has a resonance frequency around 6650 Hz, whilethe disc resonance occurs at 7320 Hz for the sixth bending mode.The wavelength for the pad mode is approximately 100 mm and forthe disc is 112 mm, considering these two bending modes, as shownin Figure 9.Figure 9. Wavelength coincidence between disc and pad.Thus, there is a wavelength coincidence between pad and disc inthis situation, leading to the appearance of a coupled mode. Figure10 shows an overlap between third pad bending mode and the sixthdisc bending mode.Figure 8. Mode shapes for brake disc.The excitation was provided by an impact hammer in the normaldirection. For this reason only the bending modes were obtained bythis modal analysis. To obtain the modal parameters in thetangential direction, a new procedure is necessary. However, thispaper addresses only modal coupling in the normal direction, i.e.,between bending modes of the disc and pads, which, for the purposeJ. of the Braz. Soc. of Mech. Sci. & Eng.Copyright 2004 by ABCMJuly-September 2004, Vol. XXVI, No. 3 / 343

M. Triches Jr et alAs a consequence, for high pressures, the third bending mode ofthe pad and the sixth bending mode of the rotor can start do vibratetogether in a coupled mode. This fact is in accordance with Figure 4,which shows noise occurrence at higher pressures.The effect of the braking pressure can be analyzed in terms ofstanding and traveling waves on the brake rotor. For the system hereanalyzed, at low pressures, rotor and pads are barely in contact. Ascontact pressure increases, the vibration amplitude peaks (on themode shape) start to move with respect to the pads. In this case,there is a traveling wave in the disc. At high pressures, rotor and padmodes start vibrating in phase. In this situation, the anti-nodes nolonger rotate, but remain in the same position along the rotordiameter, characterizing a standing wave.Behavior of Brake Components in Relation to TemperatureFigure 10. Modal coupling between modes of pad and disc.Figure 11 shows a comparison between two frequency responsefunctions, measured for rotor and pad. The third bending mode ofthe pad and the sixty bending mode of the disc are pointed out onthe graph.Figure 11. Frequency response functions for rotor and pad.Behavior of Brake Components Under PressureIt should be noticed that the resonance frequency of the twomodes involved in the potential coupling is slightly different.However, when the brake system works under pressure andtemperature, the dynamics of the brake components are changedsignificantly. This behavior is more pronounced for the pad than forthe disc, because of the friction material. The disc is made of ca

Keywords : Brake, damping, squeal, noise Introduction Disc brake noise is an ongoing problem for the automotive industry. Brake noise is perceived by customers as both annoying and an indication of a problem with the brake system. In most cases, this type of noise has little or no effect on the performance of the .

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