17 CHAPTER 5 PARAMETRIC STUDIES AND SQUEAL REDUCTION METHODS 5.1 INTRODUCTION Generally, there are a number of methods that have been used in order to reduce squeal for the improvement of passengers comfort. Structural modifications method is regarded as one of the most effective ways to reduce squeal. Structural modifications include geometric and material modifications of a disc brake assembly, removal of a part or the whole of an existing component or more often insertion of a new component (Ouyang et al 25). In this chapter influence structural modifications in order to improve squeal performance of the automotive disc brake system under evaluation, by modifying mass and stiffness distributions has been carried out. Moreover, the role of damping added to the system using pad shims has been evaluated. Theoretically, this is accomplished when the positive real parts of complex eigenvalues of the baseline model as found in the previous chapter (Figure 4.11) are reduced or eliminated. It is convenient to divide structural modifications into three general categories: material modifications, geometric modifications and adding damping shim. For the material modifications, a traditional evaluation tool has been proposed by a number of researchers to investigate the effect of material stiffness by varying Young s modulus of the disc brake components. In this research a non-traditional evaluation is introduced to examine the effects of different types of material as found in practice, which are used in
18 manufacturing disc brake components for commonly used vehicles or special heavy duty performance vehicles and racing cars. For the geometrical modifications, the first option is to examine the influence of the rotor geometry modification by changing neck thickness, neck height and top hat thickness, changing hat asymmetry, using drilled holes on the disc surface, utilizing a solid disc and altering the number of vanes. The second option is to examine the influence of modifying the pad geometry by cutting different slot and chamfer configurations, changing friction material thickness and back plate thickness. For structural modifications by inserting a new component, the addition of damping shims to the back of pad back plates is used to absorb the vibration occurred in the brake system in order to reduce the squeal occurrence. An overview of the present chapter and subsection topics is summarised in Figure 5.1. Parametric studies on squeal reduction methods Material modifications Pad shim Geometrical modifications 1. Rotor 2. Friction material 3. Caliper 4. Anchor bracket Disc Modification 1. No. of vanes 2. Hat thickness 3. Neck thickness 4. Neck height 5. Drilled rotor 6. Solid rotor 7. Hat asymmetric Pad Modification 1. Slots 2. Chamfer 3. Friction material thickness 4. Back plate thickness Reduction of squeal Figure 5.1 Overview of the chapter and subsection topics
19 5.2 MATERIAL MODIFICATIONS Material properties of the disc brake components usually have a significant effect on brake squeal generation. Researchers have reported the links between the squeal occurrence and brake components material properties. It is necessary to design the brake pads, rotor, caliper and anchor bracket such that their natural frequencies in the audible range are isolated to avoid mode coupling between them to reduce the squeal occurrence. In the assembly structure, when component s material property is altered, it leads to a significant redistribution of stiffness and mass throughout the assembly. This, in turn, will change the natural frequencies of the assembly as a whole as well as potentially varying the modes vibration. The stability of different assembly modes may be significantly altered by changing material of disc brake components. In this section, the influence of different types of materials used in manufacturing disc brake components as found in practice is examined in an attempt to select the best materials that can reduce the squeal occurrence. Ideally, the materials used in braking systems should exhibit properties, such as good thermal diffusivity and resistance to corrosion, low weight, long durability, friction stability, low wear ratio, etc. Details of the material modifications considered are given in Table 5.1.
11 Table 5.1 Proposed material modifications Component Pads Rotor Caliper Bracket Material type Young s modulus (GPa) Density (kg m -3 ) Friction material (soft).5 245 Friction material (baseline) 2.6 245 Friction material (stiff) 4 245 Cast iron (baseline) 125 7155 Al-MMC 7 28 C/C-SiC 5 21 Cast iron (baseline) 171 75 Steel 21 785 Aluminum 71 28 Cast iron (baseline) 166 75 Steel 21 785 Aluminum 71 28 5.2.1 Influence of Rotor Material The majority of rotors for conventional automotive disc brakes are made of grey cast iron (CI). Although this material is a cheap with acceptable thermal properties, sufficient mechanical strength, and satisfactory wear resistance, its density is a relatively high comparing to other materials. The weight of the rotors also increases the overall weight of the vehicle. Generally, excess weight of these parts makes handling difficult and increases fuel economy. Thus much effort has been made to utilize lightweight materials, such as aluminum metal matrix composites (Al-MMCs), carbonreinforced ceramic matrix composites (CMCs) and carbon-carbon composites, in place of CI.
111 Recently, ceramic matrix composites are considered for a high performance brake disc in automobile industry as an alternative of the conventional cast iron disc due to excellent thermo-mechanical properties as well as high strength to weight ratio. C/C-SiC is a carbon fiber phase added to a silicon carbide matrix, which has increased strength with a lower density and high tribological characteristics. In this section, the influence of two types of rotor materials found in the literature, Al-MMC and C/C-SiC, have been investigated and results compared to the existing cast iron baseline model. FE modal analysis is conducted to obtain natural frequencies of the rotor for different types of materials. Table 5.2 shows that natural frequencies of the rotor are altered for the same mode shape due to change type of materials. CEA is conducted and the results are shown in Figure 5.2. It is found that using Al-MMC material reduces the number of high unstable frequencies from 7 to 4, especially at high frequency squeal. Also, C/C-SiC rotor approximately exhibits the same results, which reduces the number of unstable frequencies to 5. From the predicted results, it is concluded that using C/C-SiC or Al MMC rotor increases the stability of the brake system and shifting squeal frequencies. Table 5.2 Natural frequencies of the rotor material Baseline Al-MMC C/C-SiC Mode shape 1453 1737 1697 3225 3858 3765 562 655 599 767 8454 825 917 197 175
112 12 CI Al-MMC SiC 1 8 6 4 2 2 4 6 8 1 Figure 5.2 Effect of rotor material modifications on brake squeal 5.2.2 Influence of Pad Material Most often, stiffness of pad material can have a significant influence on the squeal propensity of the disc brake system, since changes in the pad stiffness can alter the mode coupling between the pads and rotor. Brake pads consist of friction materials which are highly filled composite materials and back plates made of steel. Friction materials are assumed to be linear and elastic. In this section, the effect of brake pad material is evaluated by varying Young s modulus of the friction material of the pads from.5 GPa to 4 GPa. These values of modulus are in the range readily attained within brake pads available in the market (Lee et al 23). The density and Poisson s ratio of these materials are assumed to be the same as the baseline model.
113 FE modal analysis is conducted to examine the role of stiffness on changing natural frequencies of the brake pad. Table 5.3 shows that natural frequencies of the pad are changed for the same mode shape due to vary stiffness of the friction material. CEA is performed and results are shown in Figure 5.3. s of the complex eigenvalues is plotted when Young s modulus of the friction material varies from.5 GPa to 4 GPa. It is found that increasing Young s modulus of the friction material is capable of eliminating positive real parts for unstable frequencies at 1472 Hz and 2339 Hz, in addition to shifting unstable frequencies at 5816 Hz, 7384 Hz and 876 Hz. On the other hand, reducing Young s modulus of the friction material to.5 GPa, result in increasing the number of unstable frequencies and the real parts, especially at high frequency squeal. Hence, it is concluded from the above results, increasing pad modulus reduces the overall number of unstable frequencies and improves the overall system stability. The most likely physical explanation for this would be reduction in pad deformation. Table 5.3 Natural frequencies of the brake pad material Baseline.5 GPa 4 GPa Mode shape 2889 2358 3133 446 3849 4748 6735 5681 7217 8976 7475 9543
114 35 3 base o.5gpa 4Gpa 25 2 15 1 5 2 4 6 8 1 Figure 5.3 Effect of pad material modifications on brake squeal 5.2.3 Influence of Caliper Material The floating caliper housing is usually made of ductile cast iron which exhibits an elastic stress-strain relationship similar to that of steel but which is limited by the gradual onset of plastic deformation. Ductile iron is used in low cost applications where the high thermal conductivity of grey iron is less important. The modulus of elasticity for ductile iron varies from 162-17 GPa. Some brake caliper materials made from aluminum with a modulus of elasticity 7 GPa or steel with a modulus of elasticity 21 GPa. In this section, the effect of brake caliper material is evaluated for three different materials: cast iron, steel and aluminum. FE modal analysis is conducted. It is found that changing material of the caliper has a significant effect on changing the natural frequencies of the assembly, as shown in Table 5.4. From the CEA analysis, it is found that the steel caliper is capable of eliminating squeal at 2339 Hz, 5816 Hz and 7384 Hz. While caliper made of aluminum material increases the real parts for unstable frequencies of 1472
115 Hz, 876 Hz and 9471 Hz as shown in Figure 5.4. From this result, it is observed that the steel caliper generates less squeal propensity than either cast iron or aluminum caliper. Table 5.4 Natural frequencies of the caliper material Baseline Aluminum Steel Mode shape 2293 2242 2427 3964 484 4138 5667 5791 5882 6587 677 6896 8221 8413 852 2 CI Al Steel 16 12 8 4 2 4 6 8 1 Figure 5.4 Effect of caliper material modifications on brake squeal
116 5.2.4 Influence of Anchor Bracket Material The anchor bracket is used for housing the caliper and pads in the brake assembly. Different types of materials of anchor bracket are available in the market, for example cast iron, aluminum and steel. Different types of bracket materials are examined through FE modal. It is observed that changing materials has a significant effect on altering the natural frequencies of the brake assembly, as shown in Table 5.5. The influence of the anchor bracket materials on squeal generation is investigated. From Figure 5.5, it is found that the steel bracket reduces the number of unstable frequencies from 7 to 4. While the aluminum bracket significantly increases the positive real parts, especially at frequencies 1472 Hz, 2773 Hz, 5816 Hz and 9471 Hz, with a new unstable frequency generating at 976 Hz. From the simulation results, it is observed that the material of the anchor bracket also significantly influences the stability of the system. Increasing bracket stiffness reduces the overall brake squeal propensity of the system. This suggests that the steel bracket could be considered to reduce squeal generation. The results agree well with the work of Dessouki et al (23). Table 5.5 Natural frequencies of the bracket material Baseline Aluminum Steel Mode shape 88 916 941 1755 1827 1876 3164 3292 3382 468 4871 53 7533 7842 854 9262 9641 992
117 12 CI Al Steel 1 8 6 4 2 2 4 6 8 1 Figure 5.5 Effect of bracket material modifications on brake squeal 5.3 GEOMETRICAL MODIFICATIONS In this section, the influence of the geometrical modifications introduced into the disc and the pads on the brake squeal generation is highlighted. 5.3.1 Rotor Geometry Modification There are two types of rotors fitted to today's vehicles, the solid and the ventilated type brake rotor. In this research, the ventilated disc brake is used which consists of rubbing surfaces (sometimes known as cheeks) and a top hat section. The rubbing surface section is the area where a tangential friction force is generated when the disc interacts with stationary brake pads to stop the moving vehicle. The rubbing surface section is a one-piece casting with cooling fins between the braking surfaces to enable air to circulate
118 between the braking surfaces and make the rubbing surfaces less sensitive to heat build-up and more resistant to fade. The top hat section is mounted to the vehicle wheel hub and is shaped like a hat in order to protect the wheel bearings from the high temperatures induced during braking action at the rotor-pad interface. The section that connects cheeks and top hat section is known as the neck. Figure 5.6 shows a model of the brake rotor to identify the main features. Figure 5.6 Model of the brake rotor to identify main features Twelve modifications in the geometry of the brake rotor are examined for reducing the propensity of squeal generation in the disc brake system under consideration. These modifications include changing neck thickness, neck height and top hat thickness, changing hat asymmetry, using drilled holes on the disc surface, utilizing a solid disc and altering the number of vanes. The purpose of these modifications is to change natural frequency and mode shape of the rotor in order to separate merging modes. This is done by modifying mass and stiffness distributions to alter modal characteristics by changing the rotor geometry. As a result the coupled modes between the rotor and other components will be decoupled and squeal is reduced. Complex eigenvalue analysis is performed on the disc with all modifications and the results are plotted. The first important observation from simulations of modifications is that the predicted unstable frequencies are
119 shifted compared to baseline frequencies for all modifications; hence the comparison in this section depends on the mode shapes of unstable frequencies for squeal evaluation. The first modification (M1) is made by increasing the neck thickness from original value 3 mm to 5 mm. It is found that such modification reduces squeal occurrence. Since the real values of first and second unstable frequencies are reduced and the other unstable frequencies are completely eliminated. The second modification (M2) is done by increasing neck thickness of the rotor from original value 3 mm to 7 mm. This also results in real values of unstable frequencies getting reduced or eliminated. Thus these modifications have a considerable effect on reducing squeal. Figure 5.7 shows the effect of rotor modifications (M1, M2) on brake squeal. 12 base M1 M2 1 8 6 4 2 2 4 6 8 1 Figure 5.7 Effect of rotor modifications (M1, M2) on brake squeal The third modification (M3) is done by increasing thickness of top hat section of the brake rotor from original value 7 mm to 9 mm. This
12 modification seems to show an unacceptable result where there are eight squeal frequencies predicted in the complex eigenvalue analysis. More squeal frequencies are predicted for modification (M4), by reducing thickness of top hat section of the brake rotor from original value of 7 mm to 5 mm. It can be concluded that modifications M3 and M4 seem to increase the propensity of squeal and hence are not acceptable. Figure 5.8 shows the effect of rotor modifications (M3, M4) on brake squeal. 12 base M3 M4 1 8 6 4 2 2 4 6 8 1 Figure 5.8 Effect of rotor modifications (M3, M4) on brake squeal Modifications (M5, M6) are made by adding mass inside the top hat section of the brake rotor with thickness 3 mm and 1 mm, respectively, to convert the disc from symmetry to asymmetry shape. As for modification M5, there are only three squeal frequencies predicted compared to seven for the baseline model. However, these three frequencies cover both low and high frequency range. The same is the case with modification M6. This indicates that these modifications could be considered for squeal reduction. Figure 5.9 shows the effect of rotor modifications (M5, M6) on brake squeal.
121 12 1 base M5 M6 8 6 4 2 2 4 6 8 1 Figure 5.9 Effect of rotor modifications (M5, M6) on brake squeal Comparing with the baseline model, modification (M7), which is made by reducing of neck height of the brake rotor from original value 32 mm to 25 mm is capable of eliminating three unstable frequencies. Also, modification (M8), which is made by increasing of neck height of the brake rotor to 39 mm, is reasonably good where the unstable frequencies are reduced to just four with small positive real values. The results indicate that these modifications made to the rotor geometry greatly reduce squeal. Figure 5.1 shows the effect of rotor modifications (M7, M8) on brake squeal. Modification (M9) is made by reducing the number of vanes of the brake rotor from 36 vanes of baseline model to 3 which results in reduction in the mass of the rotor. The results show a significant influence of the number of vanes on system instability where the unstable frequencies are reduced to three due to a reduction in the natural frequency of the disc. Modification (M1) is done by increasing the number of vanes of the brake rotor from 36 to 42. This modification shows that five unstable frequencies
122 are eliminated, while three new unstable frequencies are generated. Hence this modification is accepted for comparatively reducing squeal from seven unstable frequencies to five, as shown in Figure 5.11. 2 base M7 M8 16 12 8 4 2 4 6 8 1 Figure 5.1 Effect of rotor modifications (M7, M8) on brake squeal 12 1 base M9 M1 8 6 4 2 2 4 6 8 1 Figure 5.11 Effect of rotor modifications (M9, M1) on brake squeal
123 Modification (M11) is made by drilling original disc brake with 12 sets of 3 holes arranged equally on the disc surface. From the results, modification M11 is found to eliminate three squeal frequencies and reducing it at the frequency 8776 Hz, while two new unstable frequencies are generated at 394 and 96 Hz. This indicates that this modification (M11) reduces squeal slightly. Modification M12 is based on the solid disc by refilling the gabs between the vanes. More squeal frequencies are predicted for this case. This modification is not capable of reducing squeal and is not a good solution. The real values of the frequencies are obviously increased and new unstable frequencies are generated. This remarks that this modification is not good solution to reduce squeal. Figure 5.12 shows the effect of rotor modifications (M11, M12) on brake squeal. 2 base M11 M12 16 12 8 4 2 4 6 8 1 Figure 5.12 Effect of rotor modifications (M11, M12) on brake squeal
124 5.3.2 Pad Geometry Modification In this section, several pad modifications are explored through FE simulation to show its effects on squeal reduction. The brake pad consists of two parts: the actual friction material and a stiff back plate. The back plate is made of steel and serves to support the friction material, which is a complex composite consisting of 2 ingredients. Four geometrical modifications of the brake pad are examined which include (i) modifying the friction material surface, utilizing chamfer configurations (ii) changing back plate thickness (iii) changing friction material thickness and (iv) modifying the friction material surface, utilizing slot configurations Figure 5.13 illustrates a model of the brake pad to identify the main features. Figure 5.13 Model of the brake pad to show main features 5.3.2.1 Influence of chamfer configurations Cutting chamfer in friction material produces three changes: loss of mass, changes of contact area and contact pressure distribution between the pad and rotor. The loss of mass by the removal of friction material alters the natural frequencies and mode shapes of the brake pad which lead to decoupling modes and eliminate squeal. Reduction of friction area also produces changes in the coupling between the pad and rotor due to the change in the distribution of contact pressure.
125 In this section, the first modification is made on the brake friction material by cutting a chamfer of 4 mm and 8 mm, respectively in both side of the pad, as shown in Figure 5.13. Complex eigenvalue analysis is performed for both modifications and the results are plotted in Figure 5.14. From the results, it is found that chamfer have a significant effects on reducing squeal, especially squeal due to out-of plane at frequencies 1471 Hz, 2339 Hz, 7383 Hz and 876 Hz. However, unstable frequencies with tangential in-plane mode not affect with chamfer modification as found in unstable frequencies 2777 Hz and 9453 Hz. It is also found that chamfer has considerable effects on reduction higher frequency squeal (up to 5 khz). In general, it can be concluded that chamfer configuration can be effective in reducing squeal at out-of-plane modes, but may be not influence the in-plane modes. 14 12 baseline chamfer4 chamfer8 1 8 6 4 2 2 4 6 8 1 Figure 5.14 Effect of chamfer on brake squeal
126 5.3.2.2 Influence of back plate thickness The second modification is done by changing the original back plate thickness from original value 5 mm to ± 1.5 mm. Variation of thickness the back plate is simulated and the results are plotted in Figure 5.15. From the results, it is found that reducing back plate thickness lead to increase the number of unstable frequencies to 8, most of them have real values greater than the baseline unstable frequencies. By increase thickness by 1.5 mm number of unstable frequencies reduce to six, most of them have real values lesser than the baseline unstable frequencies. It is observed that increasing back plate thickness can reduce squeal. 14 baseline cut 1.5 add 1.5 12 1 8 6 4 2 2 4 6 8 1 Figure 5.15 Effect of back plate thickness on brake squeal 5.3.2.3 Influence of friction material thickness One of the major contributors to brake squeal is the friction material, since squeal excitation occurs at the friction interface. A friction
127 material is composed of several types of ingredients such as friction modifiers, binders, solid lubricants, fibers, fillers and water repellent materials. The third modification is done by cutting the friction material thickness from original value of 1 mm to 8 mm and 6 mm, respectively. Variation of thickness of the friction material is simulated and the predicted results are plotted in Figure 5.16. From the predicted results, it is found that reducing friction material thickness lead to increase the squeal propensity. 14 baseline cut 2 cut 4 12 1 8 6 4 2 2 4 6 8 1 Figure 5.16 Effect of friction material thickness on brake squeal 5.3.2.4 Influence of slot configurations Changing the coupling between the rotor and pad by modifying the shape of the brake pad has been found effective. Some pad modifications like cutting slots is found to have good effects on squeal generation. The main purpose of providing pad slots is to alter the pad bending modal frequencies and a channel for debris and dust exit.
128 In this section, the geometric modification is made by changing the friction surface of the brake pads by cutting centre and diagonal slots to decouple the mode shapes and shift resonant frequencies. Slot patterns are shown in Figure 5.17. From the predicted result, it is found that the friction pads with slots reduces the number of unstable frequencies, which means fewer conditions towards squeal, as shown in Figure 5.18. The centre slot has a beneficial influence on the squeal performance of the disc brake where the unstable frequencies are reduced to just 5 with small real values. This indicates that the centre slot has a good effect on improvement the squeal performance of the disc brake. For diagonal slot, Figure 5.18 shows that reduction in the number of unstable frequencies, which implies reduction of squeal occurrence. The results suggest that the effect of diagonal slot in the pad friction material plays a significant effect in reducing squeal noise. Baseline Single centre slot Single diagonal slot Figure 5.17 Slot arrangements of the brake pads
129 14 12 base center-slot diagonal-slot 1 8 6 4 2 2 4 6 8 1 Figure 5.18 Effect of slots on brake squeal 5.4 EFFECT OF PAD INSULATOR Disc brake system has various kinds of damping, such as friction damping, friction-velocity slope damping, structural damping and additional damping by means of shim. The effects of damping have been studied on lumped models while the most of the work based on FE models neglect damping for its difficulty (Fritz et al 27). Also Ouyang et al (25) indicated that damping is very difficult to determine and model, particularly for a model of many degrees of freedom. In this section, the effect of damping shim that may have a significant effect on squeal propensity will be discussed. There are different types of damping shim that are added to back plate of the pad for reducing the squeal propensity. The most common type consists of a multi-layer cover with rubber, steel and adhesive layers. In the present brake assembly, a simplified shim which has a single-layer found in actual brake assembly is investigated.
13 The main function of the shim is to provide additional damping for the brake system. In order to simulate the shim, damping is assumed to be uniformly distributed throughout the shim, and hence the commonly used Rayleigh damping formulation is used, which is a very convenient way of accounting for damping with continuous systems. Assumed damping of the shim is given by the Rayleigh Equation 5.1, where M and K are the mass and stiffness matrices respectively. The coefficients and of the Rayleigh damping matrix are proportional to mass and stiffness respectively. C M K (5.1) The values of and are not generally known directly, but can be determined from the frequency response function measurements using Equation 5.2. i i 2 2 i (5.2) where; i is the natural circular frequency of mode i and i its damping ratio. To obtain the value of natural frequencies and damping ratios for the shim, experimental modal analysis is carried with the use of miniature impact hammer (Dytran 58SL) to excite the shim, and a miniature accelerometer (Kistler 8778A5) with a mass of.29 g to measure the response. Figure 5.19 shows experimental setup to measure the damping coefficients. The coefficients and of the Rayleigh damping matrix are 8.55 sec -1 and.5 sec respectively, which is obtained by solving the Equation 5.2 using frequencies and damping ratios at first and second bending modes, as shown in Table 5.6.
131 Figure 5.19 Experimental modal analysis for shim In the FE model, the shim is created using hexahedral solid element (C3D8) which is attractive for practical 3D analysis. It has eight nodes with three translational degrees of freedom at each node. C3D8 produces more accurate results than other elements in the finite element analysis because it is linear (p = 1), with a linear strain variation displacement mode. Tetrahedral elements are also linear, but can have more discretization error because they have a constant strain. Besides being more accurate, meshes comprised of hexahedrons are easier to visualize than meshes comprised of tetrahedrons. In addition, the reaction of hexahedral elements to the application of body loads more precisely corresponds to loads under real world conditions. The eightnode hexahedral elements are therefore superior to tetrahedral elements for finite element analysis. Then, FE modal analysis is conducted and its result is validated using experimental modal analysis. It is found that a good agreement is achieved between FE and experimental results, as shown in Table 5.6. Two damping shims are added onto the back plates of FE brake assembly and the complex eignevalue analysis is performed. The simulation results are presented in Figure 5.2. It is observed that even though this type of pad shim is simplified type, it has a good effect on the stability of many unstable frequencies and eliminating the squeal. The
132 results seem to indicate that squeal frequencies at 1472 Hz, 2339 Hz and 7338 Hz can be eliminated by addition of damping, which have out-of-plane modes. In addition, the unstable frequency at 8716 Hz is reduced. The others unstable frequencies are not affected by damping which have in-plane modes. It can be concluded that a single-layer shim could reduce or eliminate out-ofplane for all squeal frequencies, while no observed effect on in-plane squeal frequencies. Table 5.6 Comparison between natural frequencies of experimental and Mode FE modal analysis, with shim Frequency (Hz) EMA Damping ratio ( ) Frequency (Hz) 1 st bending 47.4.1 5.7 2 nd bending 145.2.25 149.6 FEA Mode shape 12 Baseline shim 1 8 6 4 2 1 2 3 4 5 6 7 8 9 1 Figure 5.2 Effect of single-layer shim on brake squeal
133 5.5 CONCLUDING REMARKS Parametric studies are carried out to examine the effect of structural modifications include changing materials of the brake components, geometric modifications of the pad and rotor and insertion of pad shims into the brake assembly on the propensity of squeal occurrence. Based on the results obtained from parametric study considered in the present study, the following general design guidelines are established to reduce the propensity of squeal. 1. The simulation results showed that using C/C-SiC or Al-MMC rotor has significant role on the stability of the brake system and shifting squeal frequencies. 2. It is concluded that increasing pad Young s modulus, using steel caliper and steel bracket reducing the overall number of unstable frequencies and improves the overall system stability. 3. It is observed that there are several disc modifications namely; increasing the neck thickness, converting the neck of disc from symmetry to asymmetry shape and changing of neck height of the brake rotor are found to reduce the squeal occurrence. 4. The results show that diagonal slot and chamfering the friction material surface provides significant squeal reduction. In addition, squeal can be reduced if the back plate and friction material thickness are increased. 5. It is also observed that a single-layer shim has a good effect on the stability of many unstable frequencies and eliminating the squeal.