brake squeal

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Brake squeal: a literature review Antti Papinniemi a, *, Joseph C.S. Lai a , Jiye Zhao b , Lyndon Loader b a Acoustics & Vibration Unit, School of Aerospace & Mechanical Engineering, University College, The University of New South Wales, Australian Defence Force Academy, Canberra, ACT 2600, Australia b PBR Automotive Pty Ltd, 264 East Boundary Road, East Bentleigh, VIC 3165 Australia Received 20 November 2000; received in revised form 18 May 2001; accepted 14 June 2001 Abstract Brake squeal, which usually falls in the frequency range between 1 and 16 kHz, has been one of the most difficult concerns associated with automotive brake systems since their inception. It cau- ses customer dissatisfaction and increases warranty costs. Although substantial research has been conducted into predicting and eliminating brake squeal since the 1930s, it is still rather difficult to predict its occurrence. In this paper, the characteristics and current difficulties encountered in tackling brake squeal are first described. A review of the analytical, experimental and numerical methods used for the investigation of brake squeal is then given. Some of the challenges facing brake squeal research are outlined. # 2002 Elsevier Science Ltd. All rights reserved. 1. Introduction Brake squeal has been one of the most difficult concerns associated with auto- motive brake systems since their inception. Research into predicting and eliminating brake squeal has been conducted since the 1930s [1,2]. Initially drum brakes were studied due to their extensive use in early automotive brake systems. However, disc brake systems are used more extensively in modern vehicles and have become the focus of brake squeal research. Figs. 1 and 2 show a typical disc brake system with a ‘‘fist type’’ caliper design. A disc brake system consists of a rotor that rotates about the axis of the wheel. The caliper assembly is mounted to the vehicle suspension system through an anchor Applied Acoustics 63 (2002) 391–400 www.elsevier.com/locate/apacoust 0003-682X/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved. PII: S0003-682X(01)00043-3 * Corresponding author. E-mail address: [email protected] (A. Papinniemi).

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Page 1: Brake Squeal

Brake squeal: a literature review

Antti Papinniemia,*, Joseph C.S. Laia, Jiye Zhaob,Lyndon Loaderb

aAcoustics & Vibration Unit, School of Aerospace & Mechanical Engineering, University College,

The University of New South Wales, Australian Defence Force Academy, Canberra, ACT 2600, AustraliabPBR Automotive Pty Ltd, 264 East Boundary Road, East Bentleigh, VIC 3165 Australia

Received 20 November 2000; received in revised form 18 May 2001; accepted 14 June 2001

Abstract

Brake squeal, which usually falls in the frequency range between 1 and 16 kHz, has been one of

the most difficult concerns associated with automotive brake systems since their inception. It cau-ses customer dissatisfaction and increases warranty costs. Although substantial research has beenconducted into predicting and eliminating brake squeal since the 1930s, it is still rather difficult to

predict its occurrence. In this paper, the characteristics and current difficulties encountered intackling brake squeal are first described. A review of the analytical, experimental and numericalmethods used for the investigation of brake squeal is then given. Some of the challenges facing

brake squeal research are outlined. # 2002 Elsevier Science Ltd. All rights reserved.

1. Introduction

Brake squeal has been one of the most difficult concerns associated with auto-motive brake systems since their inception. Research into predicting and eliminatingbrake squeal has been conducted since the 1930s [1,2]. Initially drum brakes werestudied due to their extensive use in early automotive brake systems. However, discbrake systems are used more extensively in modern vehicles and have become thefocus of brake squeal research.Figs. 1 and 2 show a typical disc brake system with a ‘‘fist type’’ caliper design. A

disc brake system consists of a rotor that rotates about the axis of the wheel. Thecaliper assembly is mounted to the vehicle suspension system through an anchor

Applied Acoustics 63 (2002) 391–400

www.elsevier.com/locate/apacoust

0003-682X/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved.

PI I : S0003-682X(01 )00043 -3

* Corresponding author.

E-mail address: [email protected] (A. Papinniemi).

Page 2: Brake Squeal

bracket. The caliper housing can slide on the anchor bracket through the two pins.Brake pads with moulded friction material can also slide on the anchor bracket. Apiston can slide inside the caliper housing. When hydraulic pressure is applied, thepiston is pushed forward to press the inner pad against the rotor and in the meantime, the housing is pushed in the opposite direction to press the outer pad againstthe rotor, thereby generating a braking torque.Like all the other applications with friction interface, noise and vibration are

inherent by-products of brake application. Brake noise and vibration has been clas-sified according to its frequency as judder, groan, hum, squeal, squelch and wire

Fig. 1. A typical ‘fist’ type brake system.

Fig. 2. Schematic of a disc brake system.

392 A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400

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brush [3]. The squeal noise that is particularly annoying usually falls into a fre-quency range from 1 to 16 kHz.Brake squeal is generated by the vibration of an unstable vibrationmode of the brake

system. In this condition the brake rotor can act as a loudspeaker since it has large flatsurfaces that can readily radiate sound. The occurrence of brake squeal is a concernsince it causes significant discomfort to the vehicle occupants and leads to customerdissatisfaction and increased warranty costs. Unfortunately, the large body of researchinto brake squeal has failed to provide a complete understanding of, or the ability topredict its occurrence [1–26]. This is partly because of the complexity of themechanisms that cause brake squeal and partly because of the competitive nature ofthe automotive industry, which limits the amount of cooperative research that ispublished in the open literature.Although a comprehensive review of brake squeal was conducted by Yang and Gib-

son in 1997 [4], it was focussed to some degree on the material aspects of a brake system.The objective of this paper is to outline the characteristics and current difficultiesencountered in tackling brake squeal and to review the analytical, experimental andnumerical methods used for the investigation of brake squeal.

2. Characteristics of brake squeal

One of the biggest contributors to brake squeal is the friction material, sincesqueal excitation occurs at the friction interface, and it normally takes approxi-mately 12 months to finalise a friction material selection. This certainly makes itvery difficult to predict a priori the propensity of a brake system to squeal. Also,often in the design of a brake system, priority is given to requirements such asbraking performance, cost and ease of manufacture. The common practice for thedifferent components of a brake system to be manufactured by different suppliersfurther complicates matters. The large number of vehicles produced means that evena low squeal propensity found during initial testing of a brake system can become amajor concern once a vehicle is in production due to a much larger population size.Modifications towards the end of development phase will have two potential risks:(1) leading to production delays and increased costs to both the brake and vehiclemanufacturers and (2) leading to products not fully validated with potential fieldwarranty concern.The most significant complication in brake research is the fugitive nature of brake

squeal; that is, brake squeal can sometimes be non-repeatable. There are manypotential squeal frequencies (unstable modes) for a brake system. Each individualcomponent has its own natural modes. The number of modes for a rotor withinhuman hearing range may be up to 80. The modal frequencies and modal shapes ofthe rotor, caliper, anchor and pad will change once these parts are installed in-situ.During a brake application, these parts are dynamically coupled together resultingin a series of coupled vibration modes, which are different from the component freevibration modes. The addition of the friction coupling forces at the friction interfaceresults in the stiffness matrix for the system containing unsymmetric off-diagonal

A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400 393

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coupling terms. From the stability point of view, this coupling is considered to bethe root cause of the brake squeal. A brake system may not always squeal given the‘‘same’’ conditions. Alternatively, small variations in operating temperature, brakepressure, rotor velocity or coefficient of friction may result in differing squeal pro-pensities or frequencies. Figs. 3 and 4 show the percentage occurrence of brakesqueal obtained at PBR Automotive Pty Ltd using a Rubore drag type noisedynamometer and an AK noise matrix for various brake pressures and temperaturesrespectively. It can be seen from Fig. 3 that there is no simple relationship betweenthe percentage occurrence and frequency of the brake squeal and the brake padpressure. Similarly, the influence of temperature on both the occurrence and fre-quency of the brake squeal is quite complex (Fig. 4).Due to the above-mentioned difficulties in designing a noise free brake system,

efforts to eliminate brake squeal have largely been empirical, with problematic brakesystems treated in a case by case manner. The success of these empirical fixesdepends on the mechanism that is responsible for causing the squeal problem. Themost fundamental method of eliminating brake squeal is to reduce the coefficient offriction of the pad material [5–7]. However, this obviously reduces braking perfor-mance and is not a preferable method to employ. The use of viscoelastic material(damping material) on the back of backplate can be effective when there is sig-nificant pad bending vibration [8,9]. Changing the coupling between the pad androtor by modifying the shape of the brake pad has also been found effective [10,11].Other geometrical modifications that have been successful include modifying caliperstiffness [12,13], the caliper mounting bracket [14,15], pad attachment method [16]and rotor geometry [17,18].

Fig. 3. Variation of occurrences of brake squeal with frequency and brake pad pressure.

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3. Analysis of brake squeal

3.1. Analytical methods

The earliest research into brake squeal suggested that the variation in the frictioncoefficient with sliding velocity was the cause [19]. Not only is there a differencebetween the static and dynamic coefficient of friction, but it was thought the drop inkinetic friction with increased sliding velocity could lead to a stick-slip condition andproduce self-excited vibration. However, squeal has been shown to occur in brakesystems where the coefficient of kinetic friction is constant [20], and has led to ana-lysis of the geometrical aspects of a brake system.Spurr proposed an early sprag-slip model that describes a geometric coupling

hypothesis in 1961 [6]. Consider a strut inclined at an angle � to a sliding surface asshown in Fig. 5(a). The magnitude of the friction force is given by

F ¼�L

1� �tan�

where � is the coefficient of friction and L is the load. It can be seen that the frictionforce will approach infinity as � approaches cot �. When �=cot � the strut ‘sprags’or locks and the surface can move no further. Spurr’s sprag-slip model consisted of adouble cantilever as shown in Fig. 5(b). Here, the arm O0P is inclined at an angle �0

to a moving surface. The arm will rotate about an elastic pivot O0 as P moves underthe influence of the friction force F once the spragging angle has been reached.Eventually the moment opposing the rotation about O0 becomes so large that O00Preplaces O0P, and the inclination angle is reduced to �00. The elastic energy stored inO0 can now be released and the O0P swings in the opposite direction to the moving

Fig. 4. Variation of occurrences of brake squeal with frequency and temperature.

A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400 395

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surface. The cycle can now recommence resulting in oscillatory behaviour.Others extended this idea in an attempt to model a brake system more completely.

Jarvis and Mills used a cantilever rubbing against a rotating disc in 1963 [21], Earlesand Soar used a pin-disc model in 1971 [22], and North introduced his eight-degree offreedom model in 1972 [23]. The culmination of these efforts was a model published byMillner in 1978 [24]. Millner modelled the disc, pad and caliper as a 6 degree of free-dom, lumped parameter model and found good agreement between predicted andobserved squeal. Complex eigenvalue analysis was used to determine which configura-tions would be unstable. Parameters investigated included the coefficient of pad fric-tion, Young’s modulus of pad material, and the mass and stiffness of caliper. Squealpropensity was found to increase steeply with the coefficient of friction, but squealwould not occur below a cut off value of 0.28. He found that for a constant frictionvalue, the occurrence of squeal and squeal frequency depends on the stiffness of padmaterial (Young’s modulus). Caliper mass and stiffness also displayed distinct nar-row regions where squeal propensity was high.The common conclusions of these models are that brake squeal can be caused by

geometrically induced instabilities that do not require variations in the coefficient offriction. Since these closed form theoretical approaches cannot adequately model thecomplex interactions between components found in practical brake systems, theirapplicability has been limited. However, they do provide some good insight into themechanism of brake squeal by highlighting the physical phenomena that occur whena brake system squeals.

3.2. Experimental methods

The frequencies of a squealing brake are highly dependent on the natural fre-quencies of the brake rotor [17]. Consequently it is of fundamental importance to beable to determine the vibration modes of the rotor. Not only will an understandingof the vibration modes of the rotor help predict how a brake system may vibrate,but it is also necessary in developing countermeasures to eliminate the problem. Theexistence of in-plane modes in addition to the bending modes is a further complica-tion, and there is evidence that the in-plane modes can be the cause of some type ofbrake squeals as well as the bending modes [18].

Fig. 5. (a) Single strut rubbing against moving surface; (b) sprag-slip system.

396 A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400

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Accelerometers provide an effective tool for determining the vibration modeshapes and the forced response of a system. Fig. 6(a) shows a bending mode shapeof a typical brake rotor that has been determined experimentally. A model was cre-ated using STARMODAL software that consisted of 384 grid points over the surfaceof a brake rotor. Frequency response measurements were made with a B&K 2032 FFTanalyser using a B&K 4374 uni-axial accelerometer and a B&K 8001 impedance head.The excitation was introduced with a B&K 4810 shaker driven by a random noise sig-nal. Unfortunately, the contact mounting required for accelerometers limits their usageon rotating brake components. They can only be used for analysis of stationarybrake components making it almost impossible to determine the mode shapes of asquealing brake rotor.Optical techniques have been used more recently. In particular, double pulsed laser

holographic interferometry has been successfully applied to squealing brake systems[16,17,25,26]. This has allowed the coupled mode shapes of a complete brake systemto be determined while it is squealing. A holographic image is produced by triggeringa laser at the maximum and minimum amplitude of a vibrating object. The differencein optical path length, caused by the deformed shape of the vibrating object, createsan interference fringe pattern on a holographic plate. The mode shape can then bedetermined by interpreting the fringe pattern.The advantage of holographic interferometry is that the mode shapes of a brake

rotor can be determined while it is squealing. Included in the holographic image canbe the rotor as well as the pads, anchor bracket and caliper. The technique can beapplied to a brake system mounted on a brake dynamometer. Suspension compo-nents, such as the spindle, spring and damper, can also be included to simulate theon car performance of the brake system.An example of the value of double pulsed holography in investigating a squealing

brake was work done by Nishiwaki et al. in 1989 [17]. In the brake system that wasbeing investigated it was apparent that the mode shape of the vibrating brake rotorwas stationary with respect to the brake caliper. Hence, the mode shape is also sta-tionary with respect to the area of excitation. The rotor was modified by changingthe symmetry of the rotor about its axis of rotation. The mode shapes of the mod-ified rotor must now rotate with respect to the area of excitation, preventing therotor from vibrating in the original vibration mode.

Fig. 6. (a) Experimental bending mode shape; (b) FEA bending mode shape.

A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400 397

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3.3. Numerical methods

Finite element analysis (FEA) has been used in the analysis of brake squeal. Modalanalysis of brake components is an area where FEA can be readily applied. Fig. 6(b)shows a finite element model of a brake rotor. The model, consisting of 8700 Tet92solid elements, was developed using a commercial finite element code ANSYS 5.6.Unfortunately, the coupling between brake components leads to vibration modes thatdiffer to those found for the individual components. Therefore, the real interestamong researchers is to be able to model an entire brake system.The critical aspect in the modelling of a complete brake system is the coupling

between components, particularly the rotor/pad interface. The contact stiffness itself isadjusted using experimental results, but the more difficult aspect is to introduce thetangential friction coupling. Liles included friction coupling between rotor and pad asoff diagonal terms in the stiffness matrix and used a complex eigenvalue analysis toassess the stability of a brake system [5]. Once the model was developed, the effect ofvarying parameters such as friction coefficient, pad geometry and caliper stiffness,could be determined. Dihua and Dongying also used a similar approach to improvethe design of an anchor bracket [14]. The work of these, and other, researchers hasshown that it is possible to create models that incorporate the friction couplingbetween the rotor and the pad. However, there has been little experimental evidence toverify the accuracy of these models. They may be useful for studying the effect ofvarying parameters within the brake system, but their ability to model the importantfriction interface is limited. As small variations in operating temperature, brake pres-sure, rotor velocity or coefficient of friction may result in differing squeal propensitiesor frequencies (Figs. 3 and 4), an accurate prediction of brake squeal using numericalmethods requires an accurate determination of material properties (particularly for thefriction material) under different operating conditions. Furthermore, proper model-ling of the boundary conditions especially where the coupling between variouscomponents is important remains a challenge.

4. Challenges for the future

Presently, research into brake squeal is focused on specific brake systems or genera-tion mechanisms. The challenge for the future is to be able to develop general techni-ques and guidelines to eliminate brake squeal during the design stage. Given thecomplexity of the mechanisms that generate brake squeal, it appears that generalguidelines are some way off in the future. For the present, the reduction of squeal noisefor specific brake systems is achievable, with the additional knowledge acquired ineach case adding to the overall understanding of brake squeal.Theoretical analysis of brake systems is difficult given the complexity of the

mechanisms and the lack of an adequate model for the friction interface that causesbrake squeal. However, this should not limit the development of simplified models asvaluable insight can be gained. Understanding obtained by studying simplifiedmodels can assist in the interpretation of experimental results and the developmentof improved computational tools.

398 A. Papinniemi et al. / Applied Acoustics 63 (2002) 391–400

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The application of FEA to brake squeal appears to offer some promise. Com-mercial software packages are being continually refined with improved modellingfeatures and the friction coupling capabilities are improving. The rapid developmentin computer aided engineering systems should make it feasible to analyse everyaspect of a brake system from braking performance to vibro-acoustic analysis, thusallowing brakes to be designed with minimum propensity to squeal and desirablebraking performance.Experimental methods will still play an important role for a number of reasons.

Firstly, they offer more effective analysis tools than numerical or purely theoreticalmethods. Secondly, diagnosis of the cause of brake squeal problems can often onlybe found by experimentation. Finally, the verification of solutions to squeal pro-blems, and the applicability of FEA models, can only be achieved through experi-mental means. Ultimately the future elimination of brake squeal will be confirmedthough experimental results and the final testing of brake systems.

Acknowledgements

This study forms part of a project funded by the Australian Research Councilunder the SPIRT scheme and the industry partner is PBR Automotive Pty Ltd.

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