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Abstract This is an off-road vehicle independently developed with spare tire mounted on the tail door which can meet military and civilian requirements. In 18, km reliability test, the system problem that tail door has severe swing and knocking on tail frame occurred. To solve this problem, the tail door system is analyzed in details. The structures of tail door and body frame are optimized, while the functions of guide block and positioning pin are improved.

And buffer blocks on upside and downside of the tail door are introduced. Through CAE analysis and road test, the tail door is improved apparently and achieves the desired effect. In recent years, the spare tire of off-road vehicle is usually placed on the floor or specialized spare tire supporting equipment [1], but in order to guarantee hard style and reduce the weight of the spare tire supporting device, spare tire is directly mounted on the tail door, as shown in Fig.

Because the weight of the spare tire has a direct effect on the tail gate, the system problem of severe swing and knocking on tail frame occurred. For the problem of reliability test, through CAE analysis of the tail door and door frame stiffness, the mechanism of problem is found and sheet metal opti- mization scheme is put forward. In the solving problem process, the mistakes of pre-design phase are. Zhang et al. For the system problem of severe swing and knocking on the tail frame occurred in reliability test, the tail door area of test vehicle is inspected and following problems are found: 1.

Cracks appear on the upper corner and lower corner of tail frame, as shown in Fig. Tail door sags, as shown in Fig. The guide block of tail door is damaged, as shown in Fig. The position pin block of tail door is damaged, as shown in Fig. Strength analysis was carried out on the body with 12 standard load cases by inertial release method. The body with standard load case is analyzed as shown in Fig. The result is shown in Fig. Table 1.

Load values are shown in Table 1. The results shows that the strength of tail frame connection area is weak. It is consistent with road test results, as shown in Figs. Design stress value and allowable value are shown in Table 1. Subsequent change validation will also use this analysis method. The result is shown in Table 1. CAE analysis with vertical load of 4 g is performed and the result can meet the requirements; due to the problems occurred in the road test, the information of road condition is collected and compared with other models.

The analysis with 6 g load in vertical direction is performed as extreme working condition. Analysis method and the load position are shown in Fig. The result shows that the tail door system stiffness is weak, as shown in Table 1. Compared with B40, the X direction displacement and Y direction in Area 1 is bigger.

According to CAE analysis, stress is larger in hinge installation area, as shown in Figs. It is bigger than the yield strength of material. The results are shown in Table 1. Lack of X direction supporting along upside and downside of tail door As shown in Fig. Upper hinge Lock. Guild block Lower hinge. Because of limitation of the rear view of tail door the area of upper frame in tail door cannot be stiffened. Thus, upper frame will swing slightly to the front of vehicle, which leads to tail door knocking on the tail frame.

The guide block in tail door on body side absorbs impact energy in Z direction and X direction from tail door by its own structure 2 mm deformation, as shown in Fig. The material of the guide block on body side is ABS, and the one on tail door side is made of aluminum alloy.

In the road test, because the impact load in the X direction and Z direction applied on the guide block in tail door on. Guild block —body side Guild block —tail door side. Lack of effective supporting in the bottom of tail door is also the reason for tail door knock on the tail frame. Unreasonable design of tail door position pin assembly structure The design of position pin assembly in tail door is unreasonable and make position pin block damaged in the road test, as shown in Fig.

The damaged pin caused supporting in X direction supporting out of control on the top of tail door, as shown in Fig. The material of position pin is 45 steel, which is mounted on the tail door with the built-in screw. The position pin block material is ABS, mounted on the outer panel of the D pillar with screws. In the road test, when the tail door swings in X direction, position pin will contact with the pin block which prevents the tail door from swing.

However, the pin is a round head, the insertion value is small, and the pin and the pin block can only contact at a point, which causes stress concentration in contact area. The material strength of pin block is weak, so pin block is damaged. Thus, supporting in X direction is out of control on the top of tail door. The lack of effective support makes the door knock on the tail frame. To optimize the structure of tail frame, detailed solutions are shown in Table 1. Results are shown in Table 1.

Roof rear beam b. Upper corner of tail frame c. Optimization Solutions area a Roof rear beam Extend D pillar panel and roof rear beam, connected together with welding point; introduce the D pillar reinforced panel, connected roof rear beam and border beam together by three M6 bolt b Tail frame upper Remove carbon-dioxide arc welding on upper corner of the tail frame. D pillar panel Welding point.

Optimization of hinge is shown in Table 1. Optimization of tail door hinge installation panel structure is shown in Fig. The stress distribution by CAE analysis is shown in Fig. The local stress in the position of upper hinge slightly exceeds the yield strength of the material, so the risk is low. Feasibility of optimization depends on the results of road test, as shown in Table 1. Increase X direction supporting along upside and downside of tail door The buffer blocks on upside and downside of tail door are added for supporting in X direction of tail door, as shown in Fig. Buffer block — Upside.

The block is vulcanized on 45 steel screws and installed in bracket which is welded with inner panel of tail door by thread. It contacts body frame triangle panel to support tail door upside along the x direction effectively. It is installed in tail door inner panel by two M6 pan head socket head screws and contact Sill trim panel to support tail door downside along the X direction effectively. It not only has a strong strength and good wear-resistance, but also has elasticity to absorb impact energy in Z direction.

Tail door guide block tail door side : the basic material is cold rolled plate O8AL, metal material. It has strong strength for support tail door in Z direction. Tail guide block rubber pad: the material is black rubber TPE, which can be a very good support and vibration absorption for the tail door in X direction Fig. Optimization of the position pin structure Because the position pin is a pin with circular head, it contact the pin block at a point. In order to avoid position pin block damaged, the position pin and pin block are optimized by using block structure, as shown in Fig.

The contact between the position pin and pin block is a surface contact. The depth of contact surface is approximately 9 mm in the X direction so that the strength in X direction is improved to support the tail door. Contact surface Section A-A. Considering all the optimization mentioned above, the displacement of tail door is calculated. The analysis method and location are shown in Fig.

According to results shown in Table 1. No damage is observed for surrounding accessories. Tail door is not sagged. The problem of the tail door system is well solved. Tail door swing problem is solved by mainly considering the stiffness of tail door, hinge, tail door frame and accessory layout.

Extending D pillar panel and roof rear beam, introducing D pillar reinforcement panel, triangle panel, sill reinforcement panel and D pillar lower reinforcement panel, and optimizing overlapping structure between D pillar and sill can improve the stiffness of whole tail door.

Adding the buffer blocks on upside and downside of tail door and optimizing the structure and material of guide block and position pin are optimized from structure side and materials side can support tail door in X direction and Z direction effectively to avoid severe swing and knocking on tail door.

Wei Z Supporting equipment of spare tire on the vehicle tail door. Chinse Patent: ,,5,10 2. Vehicle frontal collision is a common kind of vehicle collision accidents. From the early contact between the vehicle and the collision object to the collapse and deformation of vehicle structure, it takes totally about ms 0. Because the time is extremely short, energy changes rapidly during the collision process, it has a great influence on the structure and motion state of the vehicle. Zang e-mail: zanghong baicmotor. Bao e-mail: baojingru baicmotor. Song et al. The car door latch system is an important safety component of car body.

According to GB The Protection of the Occupants in the Event of a Frontal Collision for Motor Vehicle, the test requires that the door is closed but unlocked, and the technical requirements concerning the car door latch system include: 1. During the test, no door shall open; 2. During the test, no locking of the locking systems of the front doors shall occur; 3. After the impact, it shall be possible, without the use of tools: to open at least one door, if there is one, per row of seats this is only applicable to the vehicles having a roof of rigid construction.

Wherein, requirement 1 is only related to the strength and resistance to inertia load of the car door latch system. Figures 2. Both have the functions of internal unlocking, external unlocking and locking. However, they differ in structures. In the hinged door latch system, the external operating mechanism is directly connected with the latch; while in the sliding door latch system, the distance between external oper- ating mechanism and the latch is long, which makes the external operating mechanism cannot be directly connected with the latch,so a switching mechanism is needed to transmit motions to the latch.

In addition, the locking function of the. Hinged door Locking switch.


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Hinges Latch Interior door handle Lock cylinder Z. Interior door handle Sliding door Locking switch Exterior door handle Latch. During the frontal collision test, both of the left and right sliding door latch systems are locked, which does not meet the requirement of GB The Protection of the Occupants in the Event of a Frontal Collision for Motor Vehicle it shall be possible to open at least one door for per row of seats. The acceleration of B pillar is an important parameter in the vehicle collision test. Because B pillar plays a supporting role in both the front doors and rear doors, the acceleration of B pillar can reflect the acceleration of vehicle doors, and then it can be used to analyze the stress of the door latch system.

During the collision, both of the maximum acceleration of the left and right B pillar are 50 g. According to the car check after the test, both of the positions of left and right sliding doors have not changed, and the metal structure of the door and the door latch system have not yet been deformed or damaged.

The lock mechanism is an independent part in the door latch system. The lock mechanism of the sliding door latch system for the mini vehicle is integrated with the switching mechanism, and the structure is shown in Fig. The lock mech- anism is composed of lock switching panel, torsion spring and connecting rod, and. Maximum acceleration of left B pillar: Maximum acceleration of right B pillar: Connecting rod Locking position Unlocking position. Torsion spring Lock switching panel Locking switch. During the collision, the lock switching panel is turned from the unlocking position to the locking position as a result of the inertia effect, which makes the left and right sliding doors locked.

This paper conducts stress analysis on lock switching panel in state of unlocking. L rod LG. L inertia Z G X. Put the known quantities into Formula 2. Through calculation and analysis, the lock mechanism of the sliding door of the mini vehicle can only endure the frontal collision acceleration of During the vehicle frontal collision process, the collision acceleration is much greater than the gravity acceleration, so the gravity of the parts can be neglected.

According to Formulas 2. Torsion spring torque Mspring; 3. According to Formula 2. Here n is a safety factor and is generally taken as 1. After adjusting the torsion spring torque, under normal condition, when the lock switch is operated by manual, the lock switching panel will only endure the gravity G, external operation force Foperation, and torsion spring torque Mspring, as shown in Fig. Here LG and Loperation respectively represent the arm of force of gravity G and the arm of force of operation force Foperation.

According to the stress analysis, it can be concluded that:. Loperation LG. In addition, with respect to the electric door locks, after increasing torsion spring torque, the motor load will be enlarged thereupon, the risk does exist, so it still needs to do the reliability test again. As shown in Fig.

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Meanwhile, it also needs to make sure that the arrangement space and the movement space of the lock switching panel can meet the requirements. After the improvement, clump weight is added at the lower part of the lock switching panel,. Before the Improvement After the Improvement. Frod Frod Lrod.

Msping Msping Lrod. G Finertia Linertia. It meets the requirement of Formula 2. It meets the requirements of automotive ergonomics. Torsion spring Lock switching panel. Install the improved left and right sliding door switch mechanisms on the testing car, and implement the frontal collision test again. Front doors and rear sliding doors can be unlocked normally, and each door has not been locked, which meets the requirements of 4. Test status is shown in Fig. Table 2. The working principle of hinged door latch system is the same as that of sliding door, and it can also be analyzed and improved in this way; 2.

The reason why the lock system is locked during the frontal collision process is that the lock mechanism in the lock system cannot endure the collision acceleration; 3. The endurance capacity of the lock mechanism to the frontal collision accel- eration is related to the torsion spring torque and the inertia moment of lock switching panel; 4. The way of increasing the torsion spring torque can improve the endurance capacity of the lock mechanism to the frontal collision acceleration, but it will greatly affect the operation force and motor load, and it needs to be subjectively evaluated and reliability test needs to be implemented again, so this plan is not recommended.

Increasing the inertia moment of lock switching panel, can improve the endurance capacity of the lock mechanism to the frontal collision acceleration, and it has little influence on operation performance, so it is recommended to adopt this plan. Li Z Vehicle frontal collision equations and applications I. Automotive Technol —11 2. Huang J Automobile body design. China Machine Press, Beijing, pp — 3. Du Z Automobile ergonomics. It is known that idle operation of vehicles wastes fuel and produces pollution, so more and more local governments set up legislations against long-time idling operation of the vehicles.

Chen et al. Idling Start-Stop technology becomes a standard functionality in passenger cars in America and Europe, and the three kinds of mainstream schemes in Fig. Figure 3. The design life of starter for commercial vehicle is about 30, times, while real life of some starters of city buses is less than 20, times. To meet the frequent stop-start requirements, a proposed life target of , times was set. There are four new features in enhanced starter in Fig.

In order to optimize the energy storage condition of battery, a new designed LIN generator for our 7.


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The rated current of the generator is A with 28 V. LIN Pro tocol Generator. The self-developed generator control strategy was integrated in 24 V battery management system BMS. SOC is calculated in three different states—open, wake-up and sleep. The start-stop strategy mainly includes three parts: start-stop states, start control, stop control as Fig. St Req. SS Switch off light off. CAN SS release shrink. The strategy considers the vehicle heat management, the power train state, the brake safety, the battery, the generator and so on.

The start-stop lamp is used for driver choose and its state is set in Fig. The start-stop system was integrated and developed following V-model pro- gram based on HIL. The V-model of start-stop system is showed in Fig. The system picture is shown in Fig. Syetem Series definition Production.

Core Parts Function needs development Tests analyzation with Cooperation validation. The intelligent Start-Stop System saves fuel mainly by three ways—idling cut-off, high effective generating and optimized battery energy management based on driving cycles; therein, the idling cut-off is the most influencing factor. To evaluate the fuel saving effect and durability, a NO. Through a remote data sampling system which delivers the CAN data of ECU and BMS back to the data server, the timely performance data of the bus running cycle is collected.

So it is easy for engineers to process and evaluate the driving data promptly. Cooperated with Changzhou City Public Transport Company, survey, test and evaluate the performance and reliability of the start-stop system for more than one year. Based on all of test data, the average fuel saving ratio of the bus on B22 route is about 3. The partial data from October 12, to December 7, and average result are showed in Table 3.

The test result shows that the idling ratio of B22 route is about The average stop time of start-stop is about The effective start-stop times is about Currently it is recommended that any stop duration of more than 10 s can get fuel saving. Based on the fuel saving ratio and the cost, it is predicted that the end-customers can collect the cost within 1 year after the product is launched on the market.

The one-year running data of the bus proves that the developed start-stop system can get 3. In addition to the city buses, the start-stop technology can also be applied to city commercial vehicles with frequent start-stop driving cycle. Because of the low cost investment of the system and fuel saving, end-customer can collect their cost investment for less than one year. In theory, it can also get the similar performance in city-application commercial trucks. Automot Eng 8 2. Automot Shanghai 10 3. Navigant research forecasts stop-start vehicles to account for Will new CAFE standards make stop-start engine technology standard equipment?

Abstract A rectilinear rear independent suspension RRIS which has the ability to remain the wheel alignment parameters WAPs invariable in theory is presented in this paper. The kinematic analysis of the RRIS indicates that the WAPs are not sensitive to the hardpoints and the compensation length is much less than the suspension travels. The comparison results indicate that the RRIS has better WAPs and smaller forces along both lateral and longitudinal directions, which could reduce the tire wear and enhance the handling ability in application.

Suspension system plays a very important role in the vehicle chassis [1]. It general consists of guidance mechanism, shock absorber and spring [2]. As a result, the guidance mechanism synthesis is the basic issue during suspension development. Minimizing the varia- tions of wheel alignment parameters WAPs is always a goal pursued by engineers. They, of course, spent a lot of time optimizing the hardpoints and rubber bushings [4—6].

However, due to the length constraints of links, traditional suspensions, for. Xiang et al. Recently, Zhao and Liu [7, 8] proposed a novel overconstrained parallel mechanism with only one translational degree of freedom and used it as a guidance mechanism. The corresponding rear suspension system is therefore called rectilinear rear independent suspension RRIS.

It has the ability to keep the WAPs invariable according to the mobility analysis [9]. The guidance mechanism is a 4-RPR parallel mechanism in theory, in which, R denotes revolute joint and P indicates the prismatic joint. According to the mobility analysis [7], one can know that the knuckle could always trace a straight line as long as the direction of revolute joint is parallel to the horizontal plane and the axis of prismatic joint is perpendicular to the axes of adjacent revolute joints.

OXYZ represents the absolute coordinate system. X-axis denotes the longitudinal direction, Y-axis is directing to the lateral and Z-axis is directing to the vertical. The guidance mechanism is symmetrical about the YOZ-plane. As Fig. As a result, the compensation velocity and acceleration of the suspension branch are obtained by the derivation of Eq. The rotational velocity and acceleration of the suspension branch are obtained by Eq.

From Eq. Substituting the geometric parameters of sets 1 and 2 into Eqs. Table 4. By comparing the results of set 1 and set 2, one can know that the geometric parameters have a distinguished influence on the kinematics. As a result, the rea- sonable geometric parameters are very important for the RRIS. According to the variation curves, the kinematic property corresponding to set 1 is better than that of set 2, as the extreme value of set 1 is smaller than that of set 2, when their travels are identical.

From Figs. Especially, the maximum compensation length of set 1 is about 20 mm, while the suspension travel is up to mm. Therefore, the probability of severe wear of the RRIS is much lower than that of sliding suspension [10]. As a result, it could be used to many vehicles with different arrangements of chassis.

Although the WAPs of the RRIS are invariable in theory, they are changing within an interval in application, due to the elasticity and clearance of components. The structure of the suspension branch prototype is illustrated in Fig. Figure 4. The components of two suspension system, such as shock absorber, wheel, and anti-roll bar and so on are identical, except the guidance mechanisms. Accordingly, the experimental results of parallel and opposite wheel travels are shown in Figs.

The hysteretic curves indicate that there are damping, friction and clearance existing in the suspension. According to Fig. But the variation range of toe angle is larger, which indicates that the vertical torsional stiffness of the RRIS might be smaller than that one of the TTLS. Due to the larger variation ranges of wheel track and base of the TTLS, the corresponding lateral and longitudinal forces are much larger than those of the RRIS. As a result, the tire wear and impact loads of the RRIS are smaller, which will lead to a longer serve life and better handling ability.

The wheel rotational angle is determined by the variation of wheel base. From Fig. Lateral Force N 40 20 0 0 0 10 20 30 40 50 0 10 20 30 40 50 0 10 20 30 40 50 Suspension Travel mm Suspension Travel mm Suspension Travel mm d 0. Wheel Track mm 0. RRIS 0. This paper focuses on investigating the kinematics and elastokinematics of a novel rectilinear suspension system. It could make the wheel trace a straight line in theory, which is the distinguished difference from traditional ones.

The kinematic analysis indicates that although the geometric parameters have an apparent influence on the kinematics of suspension branches, they have nothing to do with the kinematics of the knuckle. And a comparison experiment of TTLS was carried out at the same time. This work lays a steady foundation for the application of the RRIS in the future. Raghavan M Number and dimensional synthesis of independent suspension mechanisms. Mech Mach Theor — 2. Henning W Automotive engineering II: vertical and lateral dynamics of vehicle. University of Technology Aachen, German 3.

Smith JH An introduction to modern vehicle design. Butterworth Heinemann, United Kingdom 4. J Automob Eng — 5. Sancibrian R, Carcia P, Viadero F et al Kinematic design of double-wishbone suspension systems using a multi-objective optimization approach. Veh Syst Dyn — 6. Cheng XF, Lin YQ Multiobjective robust design of the double wishbone suspension system based on particle swarm optimization. Sci World J —7 7.

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Int J Veh Des —68 8. Mech Mach Theor — 9.

Proceedings of SAE-China Congress 2015

Zhao JS, Li LY et al The concept design and dynamics analysis of a novel vehicle suspension mechanism with invariable orientation parameters. Veh Syst Dyn — Dixon JC Suspension geometry and computation. Based on the differential flatness theory, the flatness-based controller is designed which makes it possible to obtain the vehicle input through the vehicle output and their errors. The simulation results of the double lane change test show that the proposed control method can ensure both vehicle handing stability and path tracking effect.

Compared with vehicles equipped with ESP, the flatness-based control vehicle has better performance in both handing stability and path tracking. Andreasson studied the integrated chassis control based on vehicle handling inverse model; Menhour uti- lized the flatness theory to design the controller and achieved good effect, but the.

Wang et al. This paper proposes an integral control method for the chassis control based on the flatness-based model and the new target coming out of the road information. Since the study is aimed at the coupled longitudinal and lateral control, only the planar motion stability is considered in Fig. Thus the 5DOF vehicle dynamic model is established as Eq. The lateral and longitudinal forces of front and rear tires were given as follows. To simplify the vehicle model, Eqs. Before controller designing, the control objectives shall be established.

Proceedings of SAE-China Congress Selected Papers - Ghent University Library

There are two objectives: one is keeping handing stability and the other one is keeping path tracking. So the target path shall be set in addition to the target vehicle variables which ensure the vehicle handing stability. A double lane change path with sine function was applied. It had 3. The path is shown in Fig. The mathematical function of the path is expressed in Eq. Lateral location m 3 2. Then calculate k of Audi A6 in various speeds shown in Table 5. Table 5. According to the flatness theory, if a flat output could be found, the system would be linearised and could be controlled through linear control methods [6].

So flatness-based control method is considered for the vehicle. If the output form like Eq. Temporarily set the following formula as the flat output of the system. In this way, both the system state x and input u would be expressed by the flat output. On the other hand, the vehicle system is a differential flat system. The flatness-based controller was designed which made it possible to obtain the desired vehicle inputs through the vehicle outputs and their errors.

Driver driver c. Here red line as the desired state, and blue line as the actual status. It can be seen from the vehicle side-slip angle response in Fig. Figure 5. The trajectory has certain lateral deviation but not more than 0. It can be seen from Fig. As can be seen in Fig. Flatness-based Control. A coupled integrated control method based on differential flatness theory was put forward. It could ensure vehicle handing stability and good path tracking effect. Li D, Shangqian D, Yu F Integrated vehicle chassis control based on direct yaw moment, active steering and active stabilizer.

Veh Syst Dyn 44 2 — 3. Mokhiamar O, Abe M Simultaneous optimal distribution of lateral and longitudinal tire forces for the model following control. J Dyn Syst Meas Contr — 4. Veh Syst Dyn 44 1 — 5. Contr Eng Pract — 6. Int J Contr 61 6 — 7. Yu Z Automobile theory. China Machine Press, Beijing 8. Hui X Vehicle control method research for steering stability based on annular region. Abstract For the purpose of further improvement of the fuel economy and riding comfort of a vehicle, the centrifugal pendulum vibration absorber based on the dual-mass flywheel was discussed in order to solve the torsional vibration problem of an engine.

Firstly, the mathematical model of the centrifugal pendulum vibration absorber was established. Secondly, the matching characteristics of these design parameters of the pendulum block mass and the pendulum path forms were ana- lyzed with Matlab simulation tool. In the analysis of the mass of the pendulum block, the optimal range of the mass was obtained by setting the optimized design goal. In terms of the pendulum path forms, three kinds of pendulum paths circular path, ellipse path and polynomial path were introduced and discussed to calculate the fluctuation amplitude of the angular speed of the engine.

Finally, the results show that the optimum mass of the pendulum block is about one kilogram and the polynomial path is the optimal form for the centrifugal pendulum vibration absorber. Chen e-mail: chenxiaokai The fluctuation of crankshaft torque generated by the engine is the main torsional vibration excitation source of the transmission system of a vehicle.

Thus, the traditional torsional vibration absorber in the clutch driven plate cannot meet the new demands. The advent of vibration absorber diminished this conflict on a certain degree. Since s, the dual-mass flywheel has been widely used in passenger cars and commercial vehicles [1, 2].

At the same time, the centrifugal pendulum vibration absorber has been developed for several decades, which can theoretically counteract the vibration by choosing the appropriate absorber order [3—7]. Actually, the cen- trifugal pendulum is very small because of the space limitation and high speed fluctuation of the engine. So it cannot achieve the desired damping performance only by adding centrifugal pendulum on a crankshaft or single-mass flywheel.

In , the LuK company developed a centrifugal pendulum-type dual-mass fly- wheel vibration absorber by successfully combining centrifugal pendulum vibration absorber and dual-mass flywheel vibration absorber, which further improved the per- formance of the torsional vibration absorber. At present, this vibration absorber has been in the production stage [8, 9]. However, the research on centrifugal pendulum-type dual-mass flywheel vibration absorber in China is still at its preliminary stage.

The centrifugal pendulum-type dual-mass flywheel vibration absorber is the com- bination of centrifugal pendulum vibration absorber and dual-mass flywheel which has the advantages of the two vibration absorbers above. Research shows that when the centrifugal pendulum is attached to the second flywheel of dual-mass flywheel, the centrifugal pendulum mass can be reduced to about 1 kg, which makes it possible in a limited space [9—12].

Meanwhile, there are also a variety of different centrifugal pendulum design forms [13—16]. The damper structure is as shown in Figs. The dual-mass flywheel attenuates all orders engine vibration excitation. The centrifugal pendulum absorbs main orders engine vibration excitation by appropriate adjustments, and even prevents the occurrence of large fluctuations of the engine torque, to ensure optimum ride comfort.

So, the pendulum path can be designed to be any desired shape, not necessarily for the round. In this section the motion equations of the centrifugal pendulum vibration absorber will be derived. If l is a constant, the pendulum path is a circular arc. The potential energy of the pendulum mass due to the gravity force is too small to be ignored.

Proceedings of SAE-China Congress 2014: Selected Papers

The kinetic energy of the system is:. Generally, the centrifugal pendulum angle is small. So the differential equations of the main shaft can be reduced:. By using Eqs. Then, the pendulum path will not just for a circular arc which is a general case [6]. In general, the higher harmonic order, the smaller amplitude. At the same time, although each order harmonic phase stacks when the engine ignites sequentially, only the primary harmonic is considered in the calculation.

The second-order differential equation about the pendulum forced vibration can be derived by using Eqs. The natural frequency is proportional to the steady angular velocity of the main shaft. That means the natural frequency varies with the angular velocity, which is called the velocity adaptive. The amplitude ratio of the forced vibration can be obtained:.

The main shaft will not torsionally vibrate. So, the centrifugal pendulum vibration absorber plays a good damping performance on the n-order vibration torque over the entire speed range. The mathematical model shows that the main design parameters of the centrifugal pendulum contain the pendulum block mass m and the pendulum path form l. So the Eq. So the angular velocity fluctuation curve of the main shaft was obtained by analyzing the Eq.

The dotted line indicates the angular velocity curve which the centrifugal pen- dulum mass is equal to 0; the solid line shows the angular velocity which the centrifugal pendulum mass is equal to kg. To study the effect of the pendulum block mass m on the damping performance, this paper analyzed the angular velocity fluctuation of the main shaft by introducing the parameter m in the simulation. In the same operating conditions of an engine, as shown in the Fig. The horizontal axis indicates the pendulum block mass.

And the vertical axis presents the corresponding the fluctuation amplitude of the angular velocity of the main shaft. Then, the fluctuation amplitude of the angular velocity should meet:.

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Different pendulum paths cause different damping performance. In this paper, three forms of pendulum paths: circular, elliptical, polynomial will be analyzed in turn. As it is known, if l is a constant, the pendulum path is a circular arc. Compared to no centrifugal pendulum case, the damping performance is more obvious.

The majority of the literature also uses this form for analysis. This pen- dulum path form was also used during analyzing the mass parameter matching characteristic. The results can be found in Figs. The elliptical pendulum path means that the centroid trajectory is an elliptical curve. And its function is:. Take Eq. So the angular velocity fluctuation curve of the main shaft with elliptical pendulum path was obtained by analyzing the Eq.

The ordinate is the corresponding angular velocity fluctuation amplitude of the main shaft Dh. This means the circular pendulum path is a special case of an elliptical pendulum path. When a [ , as a increase, the fluctuation amplitude curve becomes flat. Polynomial pendulum path means that the centroid trajectory is a polynomial curve. So the pendulum parameter l is designed as a form of a polynomial. Meanwhile, because the angular amplitude of the pendulum is small, higher order items will be very small effect on the pendulum length. Then, the polynomial pendulum path will change to a circular pendulum path.

So, the circular pendulum path also is a special case of a polynomial pendulum path. So k1 ; k2 are both small. Furthermore, u4 u2 , to reflect the weight of the 4th order item, the absolute value of parameter k2 is far greater than k1. So the angular velocity fluctuation curve of the main shaft with polynomial pendulum path was obtained by analyzing the Eq. Because the polynomial pendulum path contains two parameters k1 ; k2 , the simulation result is a surface, rather than a curve.

At the same time, when the parameters a, b are set. Then, the damping per- formance is the best. In the same operating state of the engine, the same block mass of the pendulum, and the different conditions of the pendulum path forms, the maximum damping per- formance of each pendulum path form can be achieved, as shown in Table 6. Table 6. From the table, it can be seen that the ultimate damping performance of the polynomial pendulum path can be achieved and have a more substantial upgrade than other two kinds, which is the best of three.

It should be noted that the parameters of the pendulum path need to be properly designed, different engine characteristics, the optimal parameter values for each pendulum path will be some differences. The centrifugal pendulum-type dual-mass flywheel vibration absorber was mainly discussed in this paper. In the given four-stroke, four-cylinder engine condition, the mass of the centrifugal pendulum block and the parameters matching characteristics of pendulum path were analyzed respectively. Jikuan Y Mechanical vibration isolation technology.

Shanghai Science and Technology Press, Shanghai 2. Noise Vib Contr 28 5 :1—5 3. Liang CH, Tung PC Application of genetic algorithms to active vibration control of a centrifugal pendulum vibration absorber. J Sound Vib 3 — 5. Ishida Y, Inoue T, Kagawa T et al Nonlinear analysis and experiments on torsional vibration of a rotor with a centrifugal pendulum vibration absorber.

Alsuwaiyan AS, Shaw SW Performance and dynamic stability of general-path centrifugal pendulum vibration absorbers. J Sound Vib 5 — 7. J Sound Vib 5 — 8. Shi C, Parker RG Modal properties and stability of centrifugal pendulum vibration absorber systems with equally spaced, identical absorbers. J Sound Vib 21 — 9. In: 7th LuK symposium, pp 5—14 In: 8th LuK symposium, pp 55—71 Chin Mech Eng 20 15 — In: SAE international, pp — China: CN Ltd Centrifugal pendulum vibration absorber.

Matsumura S, Houjoh H Applying centrifugal pendulum vibration absorber to gear system. Abstract According to deterioration of stability and comfort caused by the increase of unspring mass of wheel-drive EV, this paper integrates the hub motor and electromagnetic suspension into the wheel and lets the motor act as the vibration absorber.

Secondly, it analyses the impact of suspension key parameters on the vibration performance and offers an optional range of parameter matching. The simulation results show that the control algorithm has obvious effect on the vibration performance. Due to shortage of energy and pollution, the electric vehicle and hybrid electric vehicle are developed rapidly.

Moreover, the wheel-drive electric vehicle has four independent drive wheels. However, the wheel-drive electric vehicle is faced with great challenges in the actual operating conditions. Firstly, the hub motor has bad working conditions, and dust, water and heat dissipation are big problems. In addition, due to mounting of hub motor, the unspring mass is increased obviously, causing deterioration of. Ren and L. The torque ripple directly acts on the wheel, and it is easy to cause resonance of the front and the rear suspensions [1—3].

In this study, the active wheel integrated by the wheel hub motor and electro- magnetic suspension is designed. The hub motor is isolated as the vibration absorber to improve deterioration of vehicle stability and comfort caused by the increasing of the unspring mass.

Further, the half-car-model based on FxLMS control strategy of the active suspension is investigated to improve vibration performance. Due to the increase of the unspring mass, the performance of the vehicle comfort decreases. The negative influence caused by the increasing of unspring mass should be restrained to ensure the ride comfort, stable working conditions of the motor and the stability of the vehicle.

The motor is isolated by a set of springs and damping in the wheel and connected with the rim by the flexible coupling so the motor and the rim can have relative motion. The mass of the motor is converted into the dynamic vibration absorber so the vibration performance degradation in the vicinity of the resonance peak can be improved obviously. The structure diagram of the active wheel is as follow Fig. The amplitude frequency characteristic image of the following three can be obtained, and compared with the rigid connection of motor and rim.

As shown in Figs. The area surrounded by the curve is less than that of the rigid connection. So RMS of the body acceleration, suspension dynamic deflection and. At the human body sensitive frequency range to the vibration 4—12 Hz , the vibration response of the vibration absorber is reduced obviously and the performance of the suspension is improved. Different parameter matches have great impact on the vibration response. For this experimental car, the initial value of each key parameter and the range of the values are shown in Table 7.

So only. Table 7. The constraint is the variable range. The result after rounding is as shown in Table 7. In this research, linear motor is integrated in the wheel as the active suspension. The linear motor can provide different force to meet the requirement of complex working conditions through the change of voltage and current. The least mean square algorithm LMS is extensively applied in the vibration and noise control. FxLMS algorithm has a high correction rate, and has a strong adaptability to non-stationary response and can quickly track the structural parameters and response character- istics.

System diagram of FxLMS algorithm is shown in the following diagram [8]. The half-car model is shown in Fig. Iy is the vehicle pitch moment of inertia. The FxLMS algorithm is applied to control the front and rear active suspensions to minimize the impact of road excitation on ride comfort. Excitation is road roughness. The primary channel is the transfer function of the road excitation to the suspension deflection. The secondary channel is the transfer function of the active suspension force to the vibration signal.

Choose wheel dynamic load as reference signal. The simulation diagram is shown in Fig. The parameters of the simulation are shown in Table 7. RMS of each key index is shown in Table 7. The control effect is obvious. In this research, the vibration absorber and electromagnetic suspension are integrated in the wheel.

The model of vibration absorber is established based on the quarter car model. The FxLMS algorithm is applied to control the electromagnetic active suspen- sion. The control strategy can alleviate the stability and comfort worsening problem due to the increasing unspring mass of the wheel-drive electric vehicle. Popularization of wheel-drive electric vehicle is promoted.

Purdy JD A brief investigation into the effect on suspension motions of high unsprung mass. Nagaya G Development of an in-wheel drive with advanced dynamic-danper mechanism. JSAE Rev 24 1 — 3. Automot Eng Int 3 4. Wei T Analysis on the vertical characteristics and vibration of the motor of the wheel drive electric vehicle. Veh Eng 36 4 — 5. Liang R, Yu Z, Ning G Suppression of negative effect of vertical vibration of electric vehicle based on the vibration absorber.

Mech Des 25 1 —30 6. J Chongqing Univ 36 8 —31 7. Vib Diagn Contr 36 2 — 8. Xu G The analysis and control for the stability and ride comfort of wheel-drive electric vehicle. Using Halbach permanent magnet array, an analysis model of actuator is built. The method of increasing end thickness is employed to decrease the ripple. After optimization, the energy— regenerative ability, electromagnetic force EM force and the force ripple have been improved a lot. Suspension system plays a vital role in improving the ride comfort and handling stability of vehicle.

Nowadays, the passive suspension is widely used in vehicles. However, passive suspension is hardly applied to the different varying road con- ditions, which cause the anti-vibration property of suspension decrease. To realize the optimal suspension performance, the active suspension is utilized to realize the optimal suspension performance, which can improve the handling stability of. Peng and L. However, the main problem of the active suspension is high energy-consuming, which increases the cost. One of the most effective methods to solve this problem is to recycle the suspension vibration energy in a moving vehicle, i.

These advantages make electromagnetic linear active sus- pension become the global popular research topic. For example, an electromagnetic linear active suspension was developed by Bose in , which can reduce roll motion of vehicle body during turning corners, pitch during braking or vibration during vehicle pasting obstacles. A permanent magnet linear motor of active sus- pension has been study by Jiabin Wang [1], which can improve ride comfort a lot.

Therefore, the permanent magnet linear motor using as active suspension is feasible and is the development direction of suspension. This paper describes analytical model and FEA model of the mentioned tubular modular permanent-magnet linear actuator with Halbach magnetized magnets. Structure parameters of this actuator have been analyzed with the method of Morris sensitivity analysis.

The method of reducing force ripple has been obtained with analysis of force ripple. The result shows that the regenerative and mechanical characteristics of the optimized actuator have been improved a lot. Figure 8. In the active working condition, active suspension adjusts the active electromagnetic force by changing the coil current. Friend Reviews. To see what your friends thought of this book, please sign up. Lists with This Book. This book is not yet featured on Listopia. Community Reviews.

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