Influence of the engine mounting system on the automotive ride comfort

Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 INFLUENCE OF THE ENGINE MOUNTING SYSTEM ON THE AUTOMOTIVE RIDE COMFORT Le Van Quynh*, Hoang Anh Tan, Nguyen Khac Minh College of Technology - TNU SUMMARY Nowadays, automotive ride comfort is one of the most important performances of automobile, the research of automotive ride comfort is getting more and more important. The aim of this study is to evaluate the influence of the parameters of engine mounting system on automotive ride comfort. To achieve this goal, a 3-D vibration model for automobile with 10 DOF is established and Matlab/Simulink is used to simulate and calculate the impact factor. The the parameters of engine mount system such as stiffness and damping coefficients are analyzed respectively according to the international standard ISO 2631-1(1997-E) for the assessment of the impact of noise and vibration to human health. The results show that the damping coefficient of engine mount system has the greatest influence on automotive ride comfort when engine operates at road surface conditions in Viet Nam. This study can provide a theoretical basis for the semi-active mounting system for engine. Keywords: engine mounting system, vibration model, stiffness, damping, ride comfort INTRODUCTION* presented to analyze the effects of vehicle Engine mount is purposed to control an parameters on ride comfort[8]. excessive motion generated from powertrain The major goal of this study is to improve a system and to isolate vibration and noise to be 3-D non-linear vibration models with 10 transmitted to main system. As vibration DOF. Matlab-simulink software is applied to design of engine mount is one of the main simulate the automotive vibration model under the conditions of roads in Vietnam. The items on the phase of vehicle development, weighted r.m.s acceleration responses of the the design should be optimized considering vertical automotive body, pitch and roll various design variables and uncertainties. In angles of automotive body are chosen as recent years, in research on engine mount objective functions. The the parameters of system, there have been a lot of papers to engine mount system such as stiffness and mention aspects such as application of damping coefficients are analyzed ANSYS, A.DAM software... etc for the respectively according to the international design of engine mount system with the standard ISO 2631-1(1997-E) for the source of excitation by itself[3-5]. The torque assessment of the impact of noise and roll axis for a mounting system with non- vibration to human health[9]. proportional damping (under oscillating AUTOMOTIVE VIBRATION MODEL torque excitation) is indeed decoupled when Physical model one of the damped modes lies in the torque The arrangement of engine mount system is roll axis direction and the study has proposed choosed four mounts in this study, so a the design optimization for engine mounting vibration model engine with 6 DOF is shown system[6]. in Fig.1. Study on the effects of the vibration vehicle Many studies indicate that the vertical engine on ride comfort movement of the vehicle body, pitch and roll angles of engine body using a linear vibration model with 8 d.o.f. is have the most impact on automotive ride mentioned by references[7]. A 3-D linear comfort, so a 3-D non-linear automotive vibration vehicle model with 10 DOF was vibration model with 10 is established to evaluate the influence of the engine mounting * Tel: 0943 141653, Email: lequynhdl@yahoo.com system on ride comfort, as shown in Fig.2. 25 Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 In Fig. 2, Kij are the suspension stiffness coefficients; Cij are the suspension damping T1 coefficients; KTij are the stiffness coefficients  of tires; CTij are the damping coefficients of 1  1  tires; Kek are the stiffness coefficients of the 1 X Y1 engine mount system; Cek are the damping 1 Z1 coefficients of the engine mount system; M and Me are the sprung mass of the automobile Fig 1. Vibration model of engine mount system  and engine; mAij are the unspung mass of the 2r K 2r  axles; IX, IY, IeX, IeY are sprung moment of C 2r 2l B 2r mA2r r Ze K C 2l B Z T2r CT2r inertia about X/Y-axis; L and Le1,2 are e1 e2 q L 2r K2l T 2l Ce1 Ce2 mA2l  K e1 Ke2  1r e C CT2l wheelbase of automobile and engine; a, b are  Te e3 B e Y KT2l q e1  Ke3 1 2l e2  K 1r X e L Y Ce4 2 distance between the centre gravity of  Ke4 x 1r Z1l C 1r X V x 1 m A1r C1l b automobile body and the centre gravity of the q C T1r K 1r 1l 1l mA1l K L1r CT1l L front/rear tires; Bf and Br are distance between q a KT1l 1l B the centre gravity of automobile body and the f centre gravity of the left/right tires; Be1 and Fig 2. Vibration model of four – wheel vehicle B are distance between the centre gravity of e2 Mathematical model engine and the left/right mount system of The combined method known as the multi- engine; ξij, Zij, Z and Ze are the vertical body system theory and D'Alembert's displacements; ,  and 1, e are the angle principle are applied to set up differential deflection at the centre gravity of the equations to describe vehicle dynamics for the automobile body and engine; x1, x2 are facilitate simulation where the object is distance between the font and rear mounting separated into subsystems linked by the force system of engine and the centre gravity of and moment equations based on multi-body system theory and D'Alembert's principle is automobile body; qij are road surface used to set up force and moment equations to roughnesses; v is the speed of automobile describe automotive dynamic system. (i=1,2 and j=left, right; k=1÷4 ). Diagram of forces and torques are shown in Fig.3.   m   F  F  F  F  A1F 1l  K1 f C1 f   TK1 f TC1 f    mA1r 1r FK1r  FK1r   FTK1r  FTC1r    m   F  F   F  F   A2l 2l K 2l C 2l TK2l TC2l  m   F  F  F  F  2r 2r  K 2r C 2r   TK2r TC2r    FC1r FK2r M Z  [F  F ]  [F  F ]  2r B  Ke1 Ce1 Ke2 Ce2 r  Z mA2r  [FKe3  FCe3 ]  [FKe4  FCe4 ]  FK1l  FC1l   FK1r  FC1r  Ze q2r FK2l  Ib1 Ib2 FC2l  F  F   F  F  FTK2r  K 2l C 2l K 2r C 2r FTC2r T 2l   A2l I   a F  F  a F  F  b F  F  MIe3 Ie4 m Y  K1l C1l   K1r C1r   K 2l C 2l  FI  FCe1 FKe1  bF  F   x .[F  F ]  x .[F  F ]  FeI FTK2l  K 2r C 2r 1 Ke1 Ce1 2 Ke2 Ce2 e FKe3 Ye FTC2l q 2l   x1[FKe3  FCe3 ]  x2[FKe4  FCe4 ]  FCe3 e    Xe FCe4 FKe4 Y  B f  I X   FK1l  FC1l  FK1r  FC1r   1r FC1r 2 FK1r X   B  r F  F  F  F   V mA1r  2 K 2l C1l K 2l C 2r FC1l F b K1l  B q FTC1r  e1 F  F  F  F  1r FTK1r   Ke1 Ce1 Ke3 Ce3  mA1l 1l  2 L  B  e2 F  F  F  F  F FTK1l Ke2 Ce2 Ke4 Ce4 TC1l q a  2 1l    i 4 i 4  B M Z   F  F f e e  Kei  Cei    i 1 i 1    L L I   e1 .[F  F ]  e2 .[F  F ]   eY 1 2 Ke1 Ce1 2 Ke3 Ce3  L L e1 e2 (1)  FKe2  FCe2   [FKe4  FCe4 ]  2 2   B Fig 3. Diagram of forces and torques e1  IeX  1  .[FKe3  FCe3  FKe1  FCe1 ]   2  B e2  .[FKe4  FCe4  FKe2  FCe2 ]  2  26 Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 In Fig. 3, FKek and FCek are the spring forces of the tire, as shown in Fig.4(a). We are and the damping forces of the engine mount known that when vehicle moves on the system; FTKij and FTCij are the spring forces roughness road, the wheel’s motion in the vertical direction could be described in two and the damping forces of tires; FKij and FCij are the spring forces and the damping forces stages: compression processes (static compression and dynamic compression) and of the suspension systems; FeI, FbI and MeI1, rebounded processes (rebounded processes M , M M are forces and moments of eI2 bI1, bI2 and wheel left-off leaving processes). As is inertia about X/Y-axis of engine and shown in Fig.4(b) and Fig.4(c). automotive body. The radial spring force of the front right The general dynamic differential equation for wheel could be determined by the following the typical four-wheel vehicle is given by formula: Eq.1. KT1r .(q1r  1r )  A difficulty in establishing the dynamic FTK1r    equations of automotive system is to find the 0 (M  m ).g nonlinear properties for suspension systems when[q  (  1r A1r )]  0 1r r1 K (5) and tires which always appear two types of T1r (M  m ).g the nonlinear (nonlinear physics and when[q  (  1r A1r )]  0 1r 1r K nonlinear geometry), when vehicle moves on T1r road surfaces. Both nonlinear geometry and The radial damping force of the front right nonlinear physics are considered in this study wheel is determined by the following and these nonlinear factor could be described formula:      (M  m ).g by the nonlinear mathematic function and the C q   1r A1r  T1r  1r 1r  when[q1r  (1r  )]  0   K (6) independent module-based programming.  T1r FTC1r   (M  m ).g Then, there will be no difficulty in finding  when[q  (  1r A1r )]  0 0 1r 1r K the solution for that.  T1r  For suspension system, the spring forces of Meanwhile, the dynamic reaction force of the the suspension systems could be determined front right wheel in vertical direction is by the following formula: defined as follows: FKij  Kij Z ij  ij  (2) FT1r=FTK1r+FTC1r (7) The damping forces of the suspension Eq.(4) and Eq.(7) are very important in systems could be determined by the creating subsystems for simulating which following formula: will be presented in the following paragraph.      C .(  Z ) when (  Z )  0.3  sc ij ij ij ij ij Road surface roughness      (3) Clcij .(ij  Z ij ) when  0.3 (ij  Z ij )  0 Road surface roughness plays an important F  Cij      C . (  Z ij ) when 0  (  Z ij )  0.3 role in evaluating vehcle ride comfort. In this  srij ij ij      study, the random excitation of road surface Clrij . (ij  Z ij ) when ( ij  Z ij )  0.3 roughness is selected the road surface Meanwhile, the dynamic reaction forces of the suspension systems in vertical highway1 in Hanoi-Lang Son section which direction is defined as follows: is measured by equipment laser ARRB Fij =FKij+FCij (4) Profiler[8]. The measuring results are For tire, a quarter of automotive model is processed by Matlab 7.0 software and the selected for anlyzing the nonlinear properties processing results are shown in Fig.5. 27 Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 z1r F =K T1r (q - ) +C T1r (q - ) M1r 1r 1r 1r 1r 1r K1r C1r wheel v F =0 1r 1r 0.03 mA1r (m + M1r )g 1r q1(m) CT1r q2(m) KT1r 0.02 q1r F1r+(m 1r + M1r )g Road surface roughness a, b, 0.01 FTK1r 0 (m + M )g/K  1r 1r 1r K 1r=tg 1r  Tyre deflection -0.01 f 1r,st q 1r 1r 1r (m + M )g Height roughness of road (m) Left-off 1r 1r -0.02 process Compression process -0.03 Rebounded process 0 10 20 30 40 50 c, Time(s) Fig. 4 Road-wheel-vehicle compled system. (a) Quarter of automotive vibration model; (b)Wheel Fig 5. Random function of road surface moving on road; (c) elastic properties of radial tire Table 1. Comfort levels related to aw threshold values aWZ values Comfort level aWZ values Comfort level Less than 0.315 m.s-2 Comfortable 0.8 m.s-2 to 1.6 m.s-2 Uncomfortable 0.315m.s-2 to 0.63m.s-2 A little uncomfortable 1.25 m.s-2 to 2.5 m.s-2 Very uncomfortable 0.5m.s-2 to 1 m.s-2 Fairly uncomfortable Greater than 2 m.s-2 Extremely uncomfortable INTERNATIONAL STANDARD ISO 2631 SIMULATION AND ANALYSIS RESULTS The most widely used international standard In order to solve the nonlinear differential for whole-body vibration (WBV) is ISO equations which presented in section 2 for 2631-1:1997E. This standard defines the evaluating influence of the parameters of methods to quantify WBV in relation to engine mounting system on automotive ride human comfort and health, perception and comfort, Matlab-Simulink software is used to motion sickness. The standard has given two simulate with a specific set of parameters of a methods for evaluation human body comfort 8-seat minibus "MEFA5-Lavi-304" and health. In this study is selected a methods manufactured in Vietnam[2], the diagram of for evaluation human body comfort, vibration simulation is shown in Fig. 6. evaluation based on the basic evaluation Simulations are carried out under the method always includes measurements of the weighted root-mean-square (r.m.s) conditions of the different road surfaces, acceleration dened by: vehicle speeds and structural parameters of the vehicle to acquire the impact factors, For T 1/ 2  1 2  (8) example, the simulation results of the vertical aw   aw (t)dt T   0  acceleration of automotive body when three vertical damping values of engine mounting where, aw(t) is the weighted acceleration (translational and rotational) as a function of system conditions of 0Cek, 0.5Cek, 1Cek are time, m/s2; T is the duration of the applied and the vehicle moves on the road measurement, s. surface highway1 in Hanoi-Lang Son section condition at v=80km/h (where Cek is used to In this way, awz, aw and aw values can be calculated from formula Eq.(8) and the r.m.s. designate the vertical damping values of value of the vertical acceleration in vehicle engine mounting system[2]) which is shown would be compared with the values in Tab 1, in Fig.7. From Fig.7 shows that the vertical for indications of likely reactions to various acceleration of automotive body (az) values magnitudes of overall vibration in the public increase while Cek values reduce which makes transport. the automotive ride comfort bad. 28 Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 the stiffness conditions of Ce = [0.8÷2.6]xCek Body values and the damping conditions of Ke=[1 1.5]xKek values have been studied, where Cek Front Suspn Engine Rear Suspn and Kek is used to designate the stiffness and damping coefficients of engine mounting Front axle Rear axle system shown in reference[2]. Effects of the Engine Suspn1 Engine Suspn4 Engine Suspn2 Engine Suspn3 Front tyre Rear tyre damping coefficients of the engine mount Fornt Road Rear Road system are shown in Fig.9. The awz value Fig 6. Simulation diagram decreases as the engine mounting system 1 1C ek stiffness increases which makes the 0.5C 0.5 ek automotive ride comfort good. When the ) 0C ek -2 damping coefficient increases from 0.8Cek to 0 /(m/s z 2.4Cek, the awz value decreases by 1.4 times. a -0.5 1 1.0C -1 ek 0 10 20 30 40 50 0.8 1.5C Time/s ) ek -2 0.6 Fig7. az when the damping coefficients of engine mount system change /(m.s 0.4 wz a Effects of the stiffness coefficients of the 0.2 engine mount system 0 0 0.5 1 1.5 2 xK K /(N.m) ek The stiffness coefficients of the engine mount k system are important factor that influence the Fig 8. Effects of stiffness automotive ride comfort. To analyze its effect on awz, aw and aw values, the stiffness 0.8 1.5K ek conditions of Ke = [0.8÷2]xKek values and the ) 0.7 1K damping conditions of Ce=[ 0.5 1 1.5]xCek -2 ek m.s 0.6 values have been studied when the vehicle /( wz moves on the road surface highway1 in a 0.5 Hanoi-Lang Son section condition at 0.4 xC v=80km/h and , where Kek and Cek is used to 0.5 1 1.5 2 2.5 3 ek C /(N.m.s-2) designate the stiffness and damping e coefficients of engine mounting system shown in reference[2]. The effects of the Fig 9. Effects of damping stiffness coefficients of the engine mount CONCLUSION system are shown in Fig.8. The awz increases as the engine mounting system stiffness The 3-D non-linear vibration model for 8- increases which makes the automotive ride seat minibus "MEFA5-Lavi-304" comfort bad. When the damping coefficient manufactured in Vietnam is developed for simulating and analyzing the influence of the increases from 0.2Cek to 1.6 Cek, the awz value parameters of engine mounting system on increases by 4.2 times. However, aw and aw values increase as the Kek value decreases automotive ride comfort. The major which is the direct cause of the growth of conclusions that can be drawn from the pitch and roll vehicle vibration [10]. analysis results as follows: Effects of the damping coefficients of the (i) The damping coefficient of engine mount engine mount system system has the greatest influence on automotive The damping coefficients of the engine mount ride comfort when engine operates at road system are another important factor that surface conditions in Viet Nam. influence the automotive ride comfort. To (ii) The combination of the stiffness and analyze its effect on awz, aw and aw values, damping coefficients of engine mount system 29 Lê Văn Quỳnh và Đtg Tạp chí KHOA HỌC & CÔNG NGHỆ 139(09): 25 - 30 have implications for improving the vibration, Mechanical Systems and Signal automotive ride comfort. Processing,Vol.20, pp.1400–1409. 5. Jun Lan (2004). Multi-Body Nonlinear Analysis (iii) In order to optimize the structural for Engine Vibration Simulation, WCCM VI in parameters of vehicle engine mount system, conjunction with APCOM’04, Sept. 5-10, 2004, one can offer an engine mount system Beijing, China. structure with intelligent control, and the 6. Jae-Yeol Park, Rajendra Singh (2008). Effect of study results in the paper can therefore serve non-proportional damping on the torque roll axis as a basis for designing the control-intelligent decoupling of an engine mounting system. Journal engine mount system through controlling the of Sound and Vibration, Vol.313. pp.841–857. 7. Le Van Quynh, Zhang Jianrun, L.D Dat, N.K range of stiffness, damping and road Binh, L.V Tuan. Research on the Optimal roughness. Parameters of Suspension Systems for Improved Ride Comfort Movement of the Vehicle by REFERENCES Vibration Model with 8 D.O.F(2010). Thai 1. Du Quoc Thinh(2008), Production of mini-car Nguyen, Viet Nam, the 10th National Conference assembly between the two governments of China on Solid Mechanics, 2010, 615-621. and Vietnam and the contents of this paper is also 8. Dao Manh Hung (2005). Effect of the part of the study in the test project maximum load of vehicle on road surface, "KC.05.DA.13". Ministerial-level Reports, Hanoi, Vietnam. 2. Nguyen Tan Chinh(2010), Study on the effect 9. ISO 2631-1; Mechanical vibration and shock- Evanluation of human exposure to whole-body of engine vibration on automobile’s ride comfort, vibration, Part I: General requirements, The Master thesis, Tnu.edu.vn. International Organization for Standardization. 3.Chang Yong Song(2010). Design Optimization 10. Le Van Quynh, Zhang Jianrun, Liu Xiaobo, and Development of Vibration Analysis Program Wang Yuan, Nguyen Van Liem(2013). Influence for Engine Mount System, of heavy truck dynamic parameters on ride comfort using 3D dynamic model, Journal of 4. Zhang Junhong, Han Jun (2006). CAE process Southeast University(Natural Science Edition), to simulate and optimise engine noise and vol.43 (4), pp. 763-770. TÓM TẮT NGHIÊN CỨU ẢNH HƯỞNG HỆ THỐNG ĐỆM ĐỘNG CƠ ĐẾN ĐỘ ÊM DỊU CHUYỂN ĐỘNG CỦA Ô TÔ Lê Văn Quỳnh*, Hoàng Anh Tấn, Nguyễn Khắc Minh Trường Đại học Kỹ thuật Công nghiệp - ĐH Thái Nguyên Ngày nay độ êm dịu chuyển động của ô tô là một trọng chỉ tiêu quan trọng nhất của ô tô, vì vậy nghiên cứu độ êm dịu càng ngày càng trở nên quan trọng. Mục tiêu chính nghiên cứu này là đánh giá ảnh hưởng của thông số hệ thống đệm động cơ đến độ êm dịu của ô tô. Để đạt được mục đích đó, một mô hình dao động phi tuyến không gian của ô tô với 10 bậc tự do được thiết lập và phần mềm Matlab/Simulink được sử dụng để mô phỏng và tính toán các hệ số ảnh hưởng. Các thông số của hệ thống đệm động cơ như độ cứng và hệ số cản lần lượt được phân tích dựa vào tiêu chuẩn Quốc tế ISO 2631-1(1997-E) về đánh giá ảnh hưởng của dao động và tiếng ồn đến sức khỏe con người. Kết quả nghiên cứu chỉ ra rằng hệ số cản của đệm cách dao động động cơ có ảnh hưởng rất lớn đến độ êm dịu khi xe hoạt động trong điều kiện mặt đường quốc lộ Việt Nam. Ngoài ra, kết quả nghiên cứu cung cấp một cơ sở lý thuyết để thiết kế hệ thống đệm bán tích cực cho động cơ đốt trong. Từ khóa: Hệ thống đệm động cơ, mô hình dao động, độ cứng, hệ số cản, độ êm dịu Ngày nhận bài:20/6/2015; Ngày phản biện:06/7/2015; Ngày duyệt đăng: 30/7/2015 Phản biện khoa học: PGS.TS Vũ Ngọc Pi - Trường Đại học Kỹ thuật Công nghiệp - ĐHTN * Tel: 0943 141653, Email: lequynhdl@yahoo.com 30