Improving dynamic and tribological behaviours by means of a Mn-Cu damping alloy with grooved surface features

D.W. Wang a, J.L. Mo a, [footnoteRef:1]*, H. Ouyang b, Z.R. Zhou a [1: *Corresponding author. Tel.: +86-28-87600601; fax: +86-28-87603142.E-mail address: jlmo@swjtu.cn (J.L. Mo)]

a Tribology Research Institute, Southwest Jiaotong University, Chengdu 610031, China

b School of Engineering, University of Liverpool, Liverpool L69 3GH, UK

Abstract:

In this work, the effect of the damping component with/without individual grooved surface features on the friction-induced vibration and noise (FIVN) and surface wear performance is studied experimentally and numerically. The experimental results show that introducing a grooved damping component in the system has a significantly improved capability in suppressing the generation of FIVN. In addition, it is observed that the friction system with a grooved damping component suffers slighter wear. Numerical results show good agreement with the FIVN events observed in the experimental test. Through analyzing the deformation behaviour of damping component and the contact behaviour of the friction system during friction process, it is speculated that the deformation behaviour of damping component plays a significant role in affecting the contact pressure and FIVN behaviour. In addition, linking the vibration performance and wear evolution, the connection between damping, and vibration and wear behaviour is discovered, which can further explain why the friction system with a grooved damping component shows improved capability in suppressing the FIVN of friction system.

Key words: Friction-induced vibration and noise; Damping; Groove; Deformation; Contact pressure; Numerical analysis.

1. Introduction

Frictional contact phenomena are commonly seen in most mechanical systems which possess relative rubbing parts [1-4]. Under the action of friction force, these systems may generate intense self-excited vibration and consequently emit squeal noise to surroundings. This friction-induced vibration aggravates the wear state of the contact interface, facilitates the occurrence of fatigue and cracks in friction materials, and eventually leads to the failure of mechanical systems [1]. In addition, the unwanted squeal is irritating, and causes noise pollution. Therefore, seeking an effective approach to reduce and control the friction-induced vibration and noise (FIVN) is extremely significant [5].

During the last decades, considerable efforts have been focused on investigating the generation mechanism of FIVN [6-10], and a large amount of work has been dedicated to identifying the effect of working parameters on the generation of FIVN [11-19]. Ibrahim et al. [1] and Akay et al. [5] provided a comprehensive review of FIVN problems. Nowadays, with the development of test approaches and computer performance, the related issues of FIVN have been widely studied experimentally as well as numerically. However, the FIVN is still characterized by typical fugitive behaviour, because of its unpredictability and high sensitivity to many working parameters [20-22]. As a consequence, the generation mechanism of FIVN is still a research matter, and the critical factors for the evolution of FIVN have not been well validated.

So far a large number of methods have been proposed to reduce the FIVN [23-43]. One of the major methods is the treatment of contact surfaces, such as modifying the physical and chemical properties of contact interfaces by using traditional or modern surface treatment technology [23-30]. The proposal of this method is based on the assumption that the FIVN is originated from the unstable vibration of contact interface. However, this method cannot be universally used in all mechanical systems involving sliding with friction, due to the limited life of newly created features on the surfaces and sometimes even prohibition of modification of the contact surfaces. Therefore, an indirect and convenient approach to reduce the FIVN problem is proposed, i.e., adding on appropriate damping components into the system. In the case of brake squeal instability of vehicles, the introduction of damping component (damping shim) on the backplate of brake pads is commonly accepted to reduce and even eliminate the FIVN [44]. In the literature, the relationship between damping and FIVN has been extensively studied experimentally and numerically [31-43]. Festjens et al. [31] found that a multi-layered viscoelastic damping shim was able to suppress the eigenmodes of a brake system in a certain frequency range. Earles and Chambers [32], Kinkaid et al. [33] all proved that inappropriate use of damping would destabilize a friction system. Kang et al. [35] numerically demonstrated that the damping shim could significantly reduce the brake’s unstable in-plane vibration. Massi et al. [36-37] and Cantone et al. [38] all proved that adding a homogenous distribution of damping material on a beam subsystem was able to stabilize a beam-on-disc system, whereas a non-uniform repartition of damping would increase the propensity of the FIVN. In addition, Sinou and Jézéquel [18, 39-40] developed mathematical expressions to show how structural damping influenced the stable and unstable regions. Brunetti et al. [41] highlighted the main role of the material damping on the dynamic response of the brake system by comparing the numerical and the experimental results. Triches Jr et al. [42] proposed the modal analysis techniques to select brake dampers for reducing braking squeal. They found that the damping shim could effectively dissipate the energy and consequently reduce squeal of the brake system.

The papers discussed above all proved the strong relationship between damping and FIVN. However, the design of real damping for a given friction system to control FIVN is always regarded as a ‘cut and paste or trial and error’ procedure, where the effects of a large matrix of modifications is experimentally evaluated on a structure [42]. This approach is sometimes successful but always expensive and time consuming. Therefore, seeking a more convenient and feasible method to add damping into friction systems is worthwhile, which will contribute to finding useful ways to reduce FIVN.

Based on the hypothesis that FIVN is originated from the unstable vibration of a mechanical system excited by the contact interface, a new idea to reduce FIVN by adding a customized damping component is proposed. If the customized damping component is beneficial for improving the contact behaviour and wear status of the contact interface, the FIVN behaviour of friction system may be accordingly improved [45]. To realise this idea, an experimental and numerical investigation of the correlation between the damping component, and contact characteristic and FIVN is carried out.

In this work, the FIVN phenomenon is studied by using a small scale tribometer, in which a customized damping component with grooved surface features is introduced for modifying the contact pressure, which in turn reduces wear and the FIVN. The purpose of this work is to provide a new insight for the design of damping components to reduce the FIVN of the contact interface. Mn-Cu damping alloy components with/without parallel grooves are introduced to the tribometer system, and their effects on reducing the FIVN and improving wear status are compared with the situation of the original friction system (without adding damping component). Moreover, numerical analysis is conducted and possible physical explanations for the experimental phenomenon are provided.

2. Experimental details

2.1 Tribometer

A small-scale tribometer characterized by typical pad-on-disc friction contact is used in this study. This simplified physical system establishes a frictional contact between the pad and disc samples, which is widely used in the material wear analysis and FIVN investigation [46-48]. The schematic of the tribometer is illustrated in Fig. 1. A disc specimen is glued on the sample table, which is driven by the rotary motor at a certain velocity. A pad specimen is mounted in the pad holder, which is bolted together with a suspension structure. A 2-D force sensor is attached to the suspension to detect both the normal and tangential forces during the friction process.

For observing the development of vibration of the system during the friction process, a 3-D acceleration sensor is fixed on the pad holder, which is able to measure the vibration signals of the friction system in three orthogonal directions. Besides, a microphone is placed near the contact surface to record the noise signal. All the vibration and noise signals measured by these sensors are collected and analysed synchronously by an 8-channel data acquisition and analysis equipment. The initial contact position and the relative motion between the pad and disc samples is also illustrated in Fig. 1.

The FIVN testing parameters are set as follows: normal force of 100 N, rotating speed of 90 rad/min, testing time of 60 min. All the tests in this study are performed in strict atmospheric conditions (60%±10% RH, 25 ℃). To ensure the good repeatability of test results, each test is repeated at least four times.

Fig. 1. Schematic of the tribometer: (1) Rotary motor, (2) Sample table, (3) Disc specimen, (4) Pad specimen, (5) Pad holder, (6) 3-D acceleration sensor, (7) Suspension, (8) 2-D strain-gauge force sensor, (9) Microphone, (10) Computer, (11) Signals acquisition and analysis system.

2.2 Contact and damping samples

Disc specimens are made of cast iron material, which possess the diameter of 25 mm and thickness of 3 mm. The Young’s modulus (E) of the disc specimens is about 158 GPa. The surfaces of the disc specimens are polished to roughness of approximately 0.04 μm Ra. The rectangular-shaped pad specimens used in the tests are a kind of composite material (HR of 50~90, E≈3 GPa) with a size of 10 mm × 10 mm × 10 mm. A typical morphology of the pad surface and EDX patterns are show in Fig. 2. The pad material is usually formed by hot compaction of coarse powders, which includes many different components (typically 10~20). According to the elemental composition analyses, the components of the pad material include binder, fillers, friction additive [49].

Fig. 2. A typical SEM morphology and EDX analysis of the pad surface.

In this work, the Mn-Cu damping alloy material is selected. It should be noted that the knowledge of how Mn-Cu damping alloy reduces FIVN is still limited, though this alloy has been widely used as shock insulation cushions. The sizes of the damping components are 10 mm×10 mm×2 mm. Two types of Mn-Cu damping alloy components, i.e., with/without parallel grooves feature are investigated in this study. These two damping components are respectively introduced between the back of the pad and the inside the pad holder, to reveal their effects on the FIVN behaviour of the friction system, as illustrated in Fig. 3. For the damping component with grooves, five parallel grooves with depths of 1 mm, widths of 1 mm and lengths of 10 mm at a regular spatial interval are manufactured on the surface in contact with the inside of the pad holder. In the test, the grooves are in parallel with the circumstantial speed direction of the pad surface. For the sake of convenience in presenting experimental and theoretical results, the friction system having a damping component with smooth surface is called D-S system; similarly, the friction system having a damping component with surface grooves is called D-G system; the friction system without a damping component is referred to as the original system. During the test process, the disc sample is changed in each test. Three pad samples are used in these tests: one is used in the original system, the other two are used in the D-S system and D-G system, respectively.

Fig. 3. Two types of damping alloy components used in the tests and the corresponding friction systems.

3. Experimental results and discussion

3.1 Noise and vibration analysis in three stages

The FIVN behaviour of a friction system is known to evolve over time, it is of necessity to study its time history. Fig. 4 shows the sound pressure and tangential vibration acceleration signals for the three systems (i.e. Original system, D-S system and D-G system) in three different stages, i.e., initial stage (890 s to 895 s), middle stage (2395 s to 2400 s) and final stage (3595 s to 3600 s). During the initial stage, visible noise and intense vibration can be detected from the original system without any damping component between the pad sample and its holder, this noise is classified as squeal according to the frequency analysis which will be shown in the following section. This phenomenon indicates that the original friction system has already become unstable and generated squeal noise. In contrast, very small amplitude of vibration and noise signals are observed from both the D-S and D-G systems (Fig. 4(a-b)). When the tests proceed to the middle stage, the noise and vibration signals measured from the original system stays at a higher level, whilst FIVN phenomenon is found to occur from the D-S system, since the noise and vibration acceleration magnitudes of the D-S system show a marked increase compared with the situations in the initial stage. In contrast, the noise and vibration magnitudes of the D-G system still stay at a low level (Fig. 4(c-d)). When the tests come to the final stage, it is clearly found that the noise and vibration acceleration magnitudes for both the original system and D-S system are now at a higher level, whereas the signal amplitudes of the D-G system are still visibly lower (Fig. 4(e-f)). Therefore, adding the damping component without any surface modification in this test is able to delay the high-amplitude vibration generation of the friction system, but cannot contain the FIVN generation during the whole testing period. While introducing the damping component with surfaces grooves into the system can thoroughly suppress the FIVN generation during the whole test period.

Fig. 4. The noise and vibration signals for the three systems in the period of 895 s to 900 s (a-b), 2395 s to 2400 s (c-d) and 3595 s to 3600 s (e-f).

In order to show the effect of the two types of damping components in reducing the FIVN behaviour of the friction system, the root-mean-square (RMS) values of the noise and vibration accelerations signals in each 100 s for the three friction systems are calculated and exhibited in Fig. 5. The RMS values can be used to represent the vibration and noise level of the friction systems. It can be observed that adding these two damping components into the friction systems can significantly reduce the FIVN intensity (RMS values), as both the D-S and D-G systems show lower RMS values compared to the original system. In addition, the D-G system has the lowest RMS values among them, which indicates that adding the damping component with surface grooves in the system is the most effective method to suppress the FIVN generation.

Fig. 5. The RMS values of the noise (a), normal vibration acceleration (b) and tangential vibration acceleration (c) of the three friction systems.

The spectral analysis of vibration acceleration signals in three stages is conducted for the three friction systems, as shown in Fig. 6. For the original system, one visible dominant frequency of 1970 Hz is observed from the initial stage to the final stage, which indicates that the original friction system has a strong tendency to generate squeal instability. For the D-S system, a similar frequency with lower energy is detected until the test proceeds to a certain stage, which suggests that adding the damping component without any surface modification in the system can delay the time of squeal instability generation and significantly reduce the FIVN level. However, it is not able to thoroughly suppress the generation of FIVN in the whole test. In contrast, for the D-G system, nearly no visible dominant frequency can be detected during the whole test, which confirms that adding the damping component with surface grooves in this work can efficiently reduce and suppress FIVN generation.

Fig. 6. Time-frequency analysis of vibration acceleration of the three friction systems in different friction stages.

3.2. Wear behaviour analysis

Fig. 7 shows the worn surface morphologies of the discs surfaces of the three friction systems observed by optical microscope after tests. For the original friction system, the disc specimen surface exhibits relatively complicated wear morphologies than the other two friction systems, with a noticeable amount of surface damage and wear debris accumulation being observed in the wear track. The deep scratches shown on the disc surface suggests that visible ploughing behaviour occurs between the contact pairs during the sliding process. For the D-S system, it is found that a smaller amount of wear debris are accumulated at the disc surface, and the degree of wear on the wear track is milder compared with the situation of the original system. While for the D-G system, the disc surface shows very mild wear, with shallow scratches and ploughings parallel to the rotating direction being observed in the wear track. Moreover, no visible wear debris is accumulated in the wear area. Therefore, the above results show that the surface wear can be considerably improved by adding a damping component into the friction system, particularly a damping component with surface grooves.

Fig. 7. Optical images of disc surfaces from the original friction system (a), D-S friction system (b) and D-G (c) friction system after the test.

The wear morphologies of the discs from the three friction systems are also examined by using SEM, to further detect the wear behaviour generated at the interfaces, as shown in Fig. 8. The disc surface of the original system exhibits very complicated and serious wear: a large amount of wear debris piles up in the worn area and forms many small plateaus, in which visible material micro-exfoliation and detachments can be seen. The EDX analysis of point A revels a higher content of O-element and Mn-element and some other pad material elements, which indicates material transformation from the pad specimen to the lower disc specimen during sliding process. As a consequence, the main wear mechanisms for the original system are abrasive wear, adhesive wear and oxidative wear. For the D-S system, although the wear level is slighter than the original system, small plateaus are still apparent on the disc surface, and some scattered wear particles can be seen as well. While for the D-G system, the scattered wear particles are not piled up and no visible plateaus can be seen on the disc surface, which indicates that the contact interface of the D-G system suffers the slightest wear among them.

Furthermore, the surface wear state can significantly affect the FIVN behaviour of the friction systems. It has been widely reported that there are contact plateaus formed on the wear surface [49-53]. These plateaus carry the main part of friction force of the systems and their deformation influences the dynamics of the global system. Ostermeyer et al. [54] verified that the self-excited vibrations of the plateaus could lead to lateral oscillations of the friction forces of the pad surface on a macroscopic scale and consequently excited the whole system to generate vibration. Therefore, more plateaus would cause a stronger tendency of friction systems to generate FIVN. Although the contact plateaus were not permanent and would be damaged and destroyed under the surface vibration and mechanical stress, the continuous surface vibration will accordingly aggravate the wear of contact surfaces, which caused the flow of wear particles and the formation of plateaus on the contact interface in return. As a consequence, the plateaus and the vibration of system affected each other.

Therefore, linking the wear and FIVN performances of the three friction systems, it is seen that the wear level of the disc surface corresponds to the intensity of FIVN. Considering the strong relationship between the contact surfaces and FIVN, it can be speculated that adding grooved damping component can improve the wear situation of contact surfaces, which accordingly may modify the FIVN behaviour of the contact interface.

Fig. 8. SEM morphologies on disc surface of the original system (a), of the D-S system (b) and of the D-G system (c).

4. Numerical results and discussion

4.1 Finite element model

In this section, a finite element analysis is carried out by using ABAQUS 6.14 to simulate the experimental process and provide a reasonable explanation for the test phenomenon. The finite element model of the test system, which is established based on the actual geometrical sizes of the tribometer, is shown in Fig. 9(a). This model mainly contains six main components of the tribometer, and the material parameters defined for all the components are consistent with the real materials properties, as listed in Table. 1. All the components are meshed by using the C3D8I (8-node linear brick, incompatible integration) element. It is worth noting that although the damping component is in tight contact with other components (the pad and the holder), deformation still occurs during FIVN process, which would cause the variation of contact states. However, a simple contact or continuity at these two locations only allows the initial static contact pressure to be found and its subsequent variation in reality would not be reflected. To allow the flexibility of the damping component to be better modelled in this analysis, instead some equivalent spring elements are introduced between the damping component and the other two components. The contact nodes at the interface between the holder and damping component have the identical coordinates. A script written by Python code is run in ABAQUS to provide node-to-node spring elements between the holder and damping component, whose number is the same as the node number of the top surface of the damping component. Similarly, the bottom surface of the damping component is supported by using a group of springs, whose number is the same as the node number of the top surface of the pad specimen, as illustrated in Fig. 9(b). For the contact pairs, the disc surface is meshed with a coarser element and set as the master surface, due to its higher hardness. The stiffness values of both contact surfaces are determined from the displacements of the damping component at those two locations. Before the spring elements are introduced into the friction system, a normal load is applied on the surface of the force sensor, then the average displacements of the damping component on both the top and bottom surfaces are calculated. Thus the contact stiffness of both surfaces can be calculated by using the normal load values divided the displacement. For the contact between the holder and the damping component, the stiffness is 6.2*107 N/m at each contact node, and for the contact between the pad and the damping component, the stiffness value of each contact node is 6.7*107 N/m.

Fig. 9(c) shows the boundary and constraint conditions set in this model: normal load of 100 N is applied on the top surface of the force sensor to establish the contact between the pad and disc specimens. The bottom holes of the rotating table are rigidly fixed except in the rotational direction, and the rotating velocity is imposed around the y-axis.

Fig. 9. Finite element model of the tribometer (a), the connection relationships among damping component and pad and holder (b) and the corresponding load and boundary conditions (c).

Table 1. The material parameters set in the finite element model

Parts

Density (kg/m3)

Young’s modulus (GPa)

Poisson’s ratio

Damping

Force sensor

2700

60

0.27

—

Suspension (plate)

7800

172

0.305

—

Suspension (shim)

1500

0.5

0.27

—

Pad holder

7800

190

0.3

—

Pad specimen

2200

1.5

0.27

—

Disc specimen

7200

158

0.275

—

Disc table

7800

172

0.305

—

Damping component

6680

60.7

0.3

α=0.04 s-1

β=6.4e-6 s

Note: The damping ration of the damping component is approximate 3.85 % at 1920 Hz.

4.2 FIVN propensity analysis

In this section, complex eigenvalues analysis (CEA) is adopted to evaluate the FIVN propensity of the three friction systems. The real parts of eigenvalues () can well reflect the FIVN tendency of the friction system. When the real parts of eigenvalues appear positive, the system will generate unstable vibration. The corresponding imaginary parts () are the excited unstable vibration frequencies. In this work, the FIVN trend of the friction system is evaluated by using the value of the negative value of damping ratio (ξ), which is also called the effective damping ratio, and is calculated based on the real parts and imaginary parts of the complex eigenvalues as the following:

. (1)

where and represent the real and imaginary parts of eigenvalue. Therefore, when the real part of eigenvalues is positive, the damping ratio (ξ) is negative and the system will generate unstable vibration. The smaller ξ indicates the higher possibility of a friction system to generate unstable vibration. A more detailed introduction about this analysis method can be found in [55].

Fig. 10(a) plots the ξ distribution of these systems, with the friction coefficient (μ) varying from 0 to 0.7. The green-star line indicates the original system having a smooth component (the component has the same size and elastic modulus as the damping component, but it has no damping), which is called ‘reference’. For the original system, when the friction coefficient increases to a value of 0.3, a negative ξ starts to occur, which indicates that the friction system becomes unstable and the friction coefficient value of 0.3 is the threshold value to excite the unstable vibration of the original friction system. With the friction coefficient increasing to 0.45, the value of ξ exhibits a sharp decrease, suggesting that the friction system has a stronger tendency to generate squeal instability at a higher friction coefficient. For the D-S system, negative ξ also occurs when the friction coefficient increases to a larger value of 0.35. However, all the values of ξ are significantly larger than those of the original system. This phenomenon indicates that adding a ‘smooth’ damping component into the friction system is beneficial for reducing the vibration tendency of the friction system, which is similar to the phenomenon observed in the experimental tests. It is worth noting that for the reference system, the ξ values are between the original system and the D-S system. This further indicates that adding damping component into this friction system is beneficial for reducing the unstable vibration. While for the D-G system, the ξ exhibits the largest value among the three friction systems with the friction coefficient increasing from 0 to 0.7, which indicates that adding grooved damping component has the best potential in stabilizing the friction system. Therefore, the CEA analysis results verify that the damping component can help reduce the vibration level, and adding grooved damping component in the system can further suppress the FIVN, which is consistent with the vibration and noise results shown in Fig. 4. In addition, to verify that the FIVN reduction is not caused by the increase of the frequency difference among the three friction systems, the natural frequencies of the three systems, at zero friction coefficient are calculated and listed in Table 2. Visibly, adding damping components into a friction system does not significantly modify the natural frequencies of the system. This indicates that the reduction of FIVN is not caused by the modification of the frequencies of the coupling modes when adding damping components.

Fig. 10(b) plots the distribution of ξ and corresponding frequencies with the friction coefficient ranged from 0.3 to 0.5. This range is based on the experimental knowledge of the possible values of the friction coefficient between two metals rubbing against each other, which is believed to cover the range of the friction coefficient between the metal disc and the ceramic pad material in this investigation. It is found that the unstable frequency (approximate 1820-1890 Hz) predicted by the CEA is very close to the tested squeal frequency of 1920 Hz. There exists a subtle difference between experimental and numerical results, which is probably caused by the required approximations and simplifications in the finite element model, such as the bolted connection at the pad holder and suspension, which is defined as a tie constraint in the numerical analysis. Therefore, the numerical results further verify that adding grooved damping component into the friction system is an effective method to further improve the FIVN behaviour of friction systems.

Fig. 10(c) exhibits the unstable mode shapes of the three systems in the case of friction coefficient equal to 0.45. Although there is a slight displacement difference among the three unstable mode shapes, the main deformation of the modes occurs in the pad and long bar-shaped pad holder, which consists of the bending deformation of suspension and pad holder carrying pad specimen sliding on the disc surface. Therefore, the damping component will not significantly affect the mode shapes of the friction system in this study.

Fig. 10. Negative damping ratio versus friction coefficient (a), the negative damping ratio versus vibration frequency in a certain friction coefficient range (b) and unstable mode shapes of the three systems (c).

Table 2. The natural frequencies of the three friction systems (from 1500 Hz to 2000 Hz)

Original system

1513.9 Hz

1707.0 Hz

1911.6 Hz

1972.4 Hz

1973.6 Hz

D-S system

1513.9 Hz

1703.7 Hz

1891.9 Hz

1972.4 Hz

1973.6 Hz

D-G system

1513.9 Hz

1698.3 Hz

1881.5 Hz

1972.4 Hz

1973.6 Hz

4.3 Physical explanation for the experimental phenomenon

Both the experimental and numerical results suggest a better capability of the grooved damping component in reducing squeal instability. In this section, a possible physical explanation is proposed to explain the phenomenon shown above.

It has been proved that when the contact pressure value on the pad surface is low and its distribution is more uniform, the FIVN behaviour of a pad-on-disc friction system can be efficiently reduced [56-57]. Thus much progress has been made to further improve the FIVN performance by shifting the highest pressure from the leading side to achieve a more uniform pressure distribution. On this consideration, the contact pressure distributions of pad surfaces for the three friction systems are summarised below, to observe the effect of lower stiffness damping components in modifying the contact pressure distributions.

The contact pressure distribution is calculated by using nonlinear static analysis, which can show how it varies at different contact positions for different friction systems. At first, a normal load is applied on the top surface of force sensor, then the disc is rotating with a predefined velocity-time curve. ‘Hard contact’ is defined to simulate contact. Fig. 11 provides the local contact pressure of the three systems at three different time steps. It is clearly observed that for the original system, the contact pressure is mainly concentrated at the leading edge and the highest pressure (dark green region) occurs at the leading point, which suggests that this friction system has a strong tendency to generate FIVN in this condition. While for the D-S and D-G systems, the contact pressures distributions are remarkably different from the original system. Apparently, the contact pressures of these two systems are more uniform and the highest pressure of pad surface varies its location constantly and does not always occur at a certain position. This phenomenon indicates that adding the damping component in the friction system is able to disrupt the contact pressure concentration at the leading edge. As shown by the authors [29], when the contact pressure was concentrated at the leading side (leading point), the friction system generated sustainable FIV; but when the highest concentrated pressure left the leading point and varied its location with time, the vibration amplitude was significantly reduced. Therefore, adding damping component can improve pressure distribution of the contact interface, prevent the pressure from concentrating at the leading point, and consequently create a favourable contact pressure distribution to reduce the unstable vibration level.

Fig. 11. Contact pressure distribution for the three friction systems at the velocity of 9.42 rad/s and friction coefficient of 0.45

Furthermore, the deformation behaviour of the damping components is exhibited to explain why the D-S and D-G systems can improve the contact pressure distribution, as shown in Fig. 12. It is observed that the damping component and the spring elements used to connect the damping component and pad specimen will deform during the sliding process. Thus, it can be speculated that this deformation will affect the contact pressure distribution of the pad surface during the friction process, which causes the variation of contact pressure distribution during the sliding process, instead of being concentrated on one particular location. Fig. 13 plots the displacements at three observation points (left (A), middle (B) and right (C) on the pad surface) in both the circumstantial (U1) and radial (U3) directions for both the D-S and D-G systems. It can be found that the deformation magnitude of the damping component of the D-G system show a larger value compared with the D-S system, regardless of the location of the observation points. Therefore, the larger displacement of the grooved damping component may modify the pad pressure distribution more significantly, and has a better capability in reducing FIVN compared with the smooth damping component.

Fig. 12. The deformation of the damping component and spring elements of D-S system (a) and D-G system (b) at the velocity of 9.42 rad/s and friction coefficient of 0.45.

Fig. 13. The schematic of the observation points in the damping component (a), the deformation displacements of point A (b), middle point B (c) and right point C (d) for both the D-S and D-G systems in the normal and tangential directions.

Combining the above experimental and numerical analysis, the inherent relation of how the damping component modifies the contact characteristics and affect the FIVN behaviour of friction systems is summarized, as shown in Fig. 14. Adding a compliant component into the system can improve the contact pressure distribution and consequently improve the FIVN behaviour of the system. Moreover, the modification of contact pressure improves the wear situation and reduces the number of contact plateaus, which can also help hinder FIVN. The lower-level FIVN will also lead to a lower number of and smaller contact plateaus. Therefore, in this work, adding more compliant components into the system can help improve the FIVN performance of friction system, and the grooved damping component possesses a further improved capability in doing this.

Fig. 14. The inherent relation of damping and FIVN and wear situation.

5. Conclusions

In this study, the ability of grooved damping alloy component on the FIVN and wear behaviour is studied by using both the experimental and numerical analysis, and the inherent relationship among damping and FIVN and wear situation are discussed. The main conclusions can be summarised as follows:

(1) Both the smooth damping and grooved damping can help reduce the FIVN of friction system. Moreover, it is found that the friction system with grooved damping component has a better performance in suppressing the generation of FIVN. The wear topographies analysis results indicate adding damping component into the system can alleviate the wear level of friction system, and the friction system with grooved damping component suffers the slightest wear situation.

(2) Numerical analysis obtains the similar results compared to the test results. It is found that the deformation of damping can reduce the pressure concentration on the pad surface, and beneficial for the pad surface to exhibit a more uniform pressure distribution. This can help reduce the FIVN level of friction system. Because the deformation of grooved damping is larger than that of the smooth damping, thus friction system with grooved damping component has a better potential in reducing the FIVN.

(3) The relationship among the damping component, FIVN and wear behaviour are summarized. The improvement of contact pressure distribution caused by adding damping component can improve the wear status of interface. The improvement of the interface wear status can also suppress the deterioration of FIVN. Meanwhile, the reduction of FIVN behaviour can also alleviate the wear level of interface. According to this analysis, it can be explained why adding the grooved damping component in system has the best ability in reducing the FIVN.

This work provides a new method to improve FIVN by using damping component. The approach used in this work can be applied in real mechanical system, such as adding the grooved damping shim into a real brake system. However, based on the limited tests conducted in this work, it is difficult to conclude the patterns of grooves used in this work can be suitable for any applicants. The further research work is to detect the ability of the damping components with different surface patterns in reducing the FIVN under various working conditions.

6. Acknowledgements

The authors are grateful for the financial support of the National Natural Science Foundation of China (No. 51675448, No. 51375408). This work is also partly supported by the National Natural Science Foundation of China (No. 11672052).

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