Pad formulation impact on automotive brake squeal

Transcription

1 Pad formulation impact on automotive brake squeal L. MORTELETTE a, b, J-F. BRUNEL a, X. BOIDIN a, Y. DESPLANQUES a, P. DUFRENOY a, L. SMEETS b a. Laboratoire de Mécanique de Lille, UMR CNRS 8107, Avenue Paul Langevin, Villeneuve d Ascq Cedex, France. b. Lapinus Fibres bv, Delfstoffenweg 2, 6045 JH Roermond, The Netherlands Abstract : Specific friction materials have been developed and then tested on a braking Tribometer in order to study the impact of mineral fibres on the NVH behaviour of the braking pads. Two parameters were tested: fibres amount and fibres type. Numerical simulations have been used for determining instability conditions. Results show that the squealing mode is selected among instable modes. That selection depends on the type and the amount of mineral fibres in the friction material. Rubbing surfaces have been observed using SEM. Silent and noisy surfaces present very different morphologies. Résumé : Des matériaux de friction spécifiques ont été développés puis testés sur tribomètre de freinage afin d étudier l influence du type et du taux de fibres minérales sur le comportement vibroacoustique des garnitures de frein. Des simulations numériques ont été menées pour déterminer les conditions d instabilités. Les résultats montrent une sélection du mode crissant parmi ceux instables en fonction du type et du taux de fibres minérales présentes dans la garniture de frein. Les surfaces de frottement, observées au MEB, ont montré des faciès très différents en fonction de leur comportement crissant ou silencieux. Key words: squeal noise, mineral fibres, instabilities, contact interface 1 Introduction Most of the commonly car braking systems generate braking force by forcing friction materials against a rotating disc. Sometimes squeal noise may be generated but it is always unwanted. In such phenomenon, the sound pressure level is greater than 80dB in a frequency range between 1 and 16 khz. Squeal noise has been found to come from the unstable behaviour of the system due to friction induced vibrations corresponding to one or few natural frequencies of the braking system [1]. Among squeal generation mechanisms, the most important mechanism leading to self-excited vibration is the mode coupling. Regarding to contact conditions, by means of experimental investigation on real disc brake system, it can be established that the friction pad strongly influences the squeal occurrences, especially brake pad material formulation. The objective of this paper is focused on the impact of the use of mineral fibres in friction material on NVH (noise, vibration and harshness) behaviour of disc brake system and two parameters have been particularly studied: the amount of fibres and the type of fibres. The methodology of this study is based on experimental and numerical approach with simplified formulations. Friction materials have been prepared following similar process as used in industry. Tests on Tribometer are used to characterise friction behaviour and dynamical response of the contact, with vibratory and acoustic measurements. Rubbing surfaces have been observed (SEM) after test in order to characterise the third body and to identify the role of mineral fibres on friction mechanisms activated in contact. In addition a numerical model of this set-up based on complex eigenvalues analysis has been performed to determine the unstable modes and to give additional data for understanding. 1

2 2 Experimental set-up Stop braking tests have been performed on the LML tribometer device [2]. The device is composed by an electric motor with a power of 11.5kW; a transmission beam equipped with a flywheel and a torquemeter and by a test chamber. Tests on tribometer have been performed in order to characterise the NVH behaviour of the studied friction materials using instrumentation permitting to perform high frequency dynamic and acoustic measurements: 3D accelerometers on the pad, piezoelectric load sensor at the rear of the pad, displacement laser sensor and microphone. 2.1 Friction materials Six formulations have been developed for the study: the aim is to focus on the role of the mineral fibres. Those formulations are simpler than commercial ones because they contain only six components. All of the formulations contain the same common base (80vol %) and only 20vol% is different: it can be 20vol% of mineral fibres, 20vol% of barites or 10vol% of mineral fibres and 10vol% of barites. Three of those formulations were used to show the impact of the amount of fibres and designations for these formulations are: A0 (no fibres), A10 (10vol%fibres) and A20 (20vol%fibres); four of them were used to show the impact of the fibre type (all containing 10vol% of mineral fibres) and designations for these formulations are: A10, B10, C10 and D10. The friction materials have been prepared following classical industrial methods. The table 1 summarises the composition of the 6 used formulations. Compression modulus measurements of those formulations have been performed on a classical electromechanical compression device, showing slightly variations between all the formulations. FIG. 1 Composition and compression modulus of the formulations developed for the study. 2.2 Braking discs Discs used for the braking tests are made of grey cast iron, for each formulation, a new disc is used. Preparation is composed by 2 steps: machining and polishing, each step is completed with metrology measurements. The goal is to achieve to have the same defects in the same ranges for all the discs. Typical disc circumferential disc profile after mounting on the Tribometer is shown in Figure 2. Profile presents 2 hollows and 2 bumps with a difference around 45 µm between higher and lower point. FIG. 2- Disc profile after mounting on Tribometer. 2

3 2.3 Experimental protocol A preliminary campaign has been set up in order to build experimental protocol. The aim was to find parameters which lead to noise generation during braking tests. Each parameter range has been investigated: rotation speed, inertia and applied load. Results show that noise appears during the last rotations of the disc which means at low speed. More noise occurrence was generated using high load and high inertia which means a high dissipated energy. The following protocol has been chosen: an initial speed of 500 rpm (fixed for all the tests), an applied load of 1000 N (fixed for all the tests) and the varying parameter is inertia between 10 to 20 kg.m² (program is presented in figure 3). FIG. 3 - Inertia program used for stop braking tests: 2 successive cycles of an increase and a decrease of inertia. 3 Numerical simulations Two numerical approaches are classically adopted in squeal occurrences [3]: the non linear transient dynamic analysis [4] and complex eigenvalues analysis. In this paper, the complex eigenvalues analysis is performed because of his efficiency in time consuming compared to transient computation even if not all frequencies predicted to be unstable are always found experimentally. Analysis is decomposed in two steps: a quasi-static analysis permits to calculate the pressure distribution in the interface of contact including the normal load and the rotational velocity of the disc. The contact interface between the disc and the pad is described using an Augmented Lagrangian algorithm with 3 main parameters: the friction coefficient, the normal contact stiffness (k n =1.0E14 N/m) and the penetration tolerance. The pressure distribution is then used in the next step: the complex modal analysis permits to obtain the complex eigenvalues of the model (α+jω). Unstable modes are defined by a real part (α) different of zero. That study permits in a first time to identify those unstable modes and then it permits to study the influence of parameters on the instabilities. The numerical model (Figure 5) has been developed under ANSYS software. It is composed by the disc, the pad, the braking system and the contact interface. Disc and pads are defined by using finite elements. The finite element geometry is the same as the real geometry in terms of shapes and dimensions. The braking system is defined by three stiffness in each direction (normal, tangential and radial). Normal stiffness is given by the piezoelectric load sensor and tangential and radial stiffness are given by the steel arm of the system which give tangential and radial rigidity to the system. The stiffness of the piezoelectric load sensor is higher (k z =1.0E10 N/m, characteristic of the sensor) than steel arm stiffness (k x =k y =1.0E8 N/m, estimated in the two directions). Experimental modal analysis was performed using a shock hammer and an accelerometer. Numerical modal analysis was performed in order to find the correct boundary conditions which give close results to the reality (in fact to experimental results). The results show that the best configuration is when nodes of plane contact surface are clamped in radial direction and where nodes of the six screws are clamped in the 3 directions. Disc s finite element model is refined and adjusted to make numerical eigenvalues close to experimental modal analysis. Fig. 5 Numerical model parameters. 3

4 4 Results 4.1 Experimental results Global noise occurrence results for all the formulations according to the protocol described in the section 2.3 are given in Figure 5 Noise occurrence is computed by dividing the noisy braking by all the braking. More noise occurrences were obtained with the A10 and D10 formulations whereas the formulation B10 gives no noise and the formulation C10 gives a few noises: only two brakings have been noisy. FIG. 5 - Global results: noise occurrence summary during Tribometer tests. Spectral analyses on noise raw signals and force measurements have been performed for each noisy braking. For all the noisy brakings (all formulations and all inertia), a common excited frequency may be identified around 6700 Hz. In addition, two formulations (A20 and D10) present other excited frequencies around 3000, 6500 and 8000 Hz (Figure 6a). Results show that noise occurrences are not continuous and are mostly generated at the end of the brakings during the last rotation of the disc and so at low speed. Figure 6b presents the transient evolution of the three force components (normal, tangential and radial) and the noise raw signal for the four last disc rotations. Perturbations may be observed at the same time on all the signals Contact forces fluctuation can be correlated to the disc undulations (with 2 hollows and 2 bump as previously described in Figure 2). Squeal occurrence is characterized by dynamical perturbations on all signals. It is shown that squeal mainly appears on an increasing slope of the disc. FIG. 6 (a) Spectral analyses on noise raw signals. (b) Correlation between noise and efforts signals. 4

5 4.2 Numerical results Unstable modes are determined with a complex eigenvalues analysis. The first step is the quasi-static analysis which calculates the pressure distribution in the contact (Figure 7a). A constant and uniform friction coefficient is used for simulation purpose issued from experimental investigations (by dividing the tangential force by the normal one). Localisation in the inner radius area of the contact and overpressure in the front of the contact which is due to the speed effect may be observed. In the second step, unstable modes are determined (corresponding to a positive real part of the mode (Figure 7b)). Results show 3 unstable frequencies: at 2816Hz, 6566Hz and 9267Hz. It is important to notice that one of the unstable frequencies is equal to 6566Hz very closed to the value of the frequency of the noise which occurred during tribometer tests: 6700Hz. The two other unstable modes are also observed during the stop braking tests of two formulations (A20 and D10). The next part of the study has been focused on the study of the mode corresponding to that frequency of 6566Hz. A parametric analysis has been performed for the unstable mode of 6566Hz. Two parameters have been investigated: friction coefficient from to 1 and pad compression modulus from 3000 to 6000MPa. The evolution of the unstable complex frequency in function of the studied parameter (µ or E) has been observed. Results show that unstable modes refer to the coupling of two modes of the system: one mode of the disc (out of plane mode with 4 nodal diameters) with one mode of the pad. For a µ=0.28, the two frequencies are coupled leading to an unstable mode. The instability can be observed for a friction coefficient up to 0.9. For greater values of µ, the coupling disappears. The same behaviour can be observed for the compression modulus as illustrated in Figure 7c. (a) Domains of instability (b) FIG. 7 (a) Contact pressure distribution, (b) unstable modes, (c) influence of friction coefficient and compression modulus on unstable modes. 5 Discussion There is a strong correlation between experimental and numerical results: experimental squealing frequencies correspond to numerical unstable frequencies however not all of the unstable modes found with numerical approach are excited. Excited modes depend on the friction material and on history effect. Moreover, the parametric analysis permits to identify instability domains and so parameter ranges within it is possible to have instabilities and thus noise occurrences. The ranges are between 0.28 and 0.9 for friction coefficient and between 3000 and 5500MPa. It is important to notice that those ranges entirely cover experimental results: tribometer tests show that friction coefficient is comprised between 0.3 and 0.5 and friction materials characterisation results show that pad compression modulus is comprised between 3000 and 5460MPa. That means that all the brakings should lead to noise generation but this is not the case. This may be due to the simple contact interface of the model which includes uniform friction coefficient and uniform normal contact stiffness whereas rubbing surfaces observed using SEM device[5] show great differences between noisy and silent brakings[6] (see Figure 8). 5 (c)

6 FIG.8 SEM observation of contact surfaces: silent braking (left) and noisy braking (right). 6 Conclusion NVH behaviour of friction materials have been studied and characterised using numerical approach and experimental tests focusing on the impact of mineral fibres in formulations. Experimental investigations on braking tribometer permit to characterise braking squealing noise. A common frequency has been identified in all the noisy brakings spectrums: 6700Hz. In addition, correlations between noise occurrences, efforts and disc undulations have been established. A simplified model of the tribometer has been developed to interpret the experimental results. Numerical model based on a complex eigenvalues analyses also gives an unstable frequency of 6566Hz, which corresponds to the frequency experimentally obtained during tribometer tests. It permits then to identify the type of coupling which lead to that unstable frequency. A parametric analysis has been performed to show the impact of friction coefficient and compression modulus on the coupling of frequencies. It has been shown that all the formulations should lead to squeal at 6.7 khz; however this squeal frequency was only experimentally obtained for 4 friction materials and not for all the stop braking. There is a need to investigate more precisely and to improve the contact interface in order to study the establishment of instabilities leading to noise generation. References [1] Akay A., Acoustics of friction, Journal of Acoustical Society of America 111, , 2002 [2] Desplanques Y., Degallaix G., Copin R., Berthier Y., A tribometer for the study of materials under railway braking conditions, Tribology and Interface Engeneering Series 39, , 2001 [3] Ouyang H., Nack W., Yuan Y., Chen F., Numerical analysis of automotive disc brake squeal: a review, International Journal of Vehicle Noise and Vibrations 1, , 2005 [4] Brunel J-F., Dufrénoy P., Transient Analysis of Squealing Mode Selection in Disc Brake, SAE International, 49-52, 2008 [5] Massi F., Berthier Y., Baillet L., Contact surface topography and system dynamic of brake squeal, Wear 265, , 2008 [6] Eriksson M., Bergman F., Jacobson S., Surface characteristic of brake pads after running under silent and squealing conditions, Wear 232, ,