裝配圖大學生方程式賽車設計（整體車架、標準安全系統(tǒng)、座椅及附件設計）（有cad圖+三維圖+中英文翻譯）,裝配,大學生,方程式賽車,設計,整體,總體,車架,標準,全系統(tǒng),座椅,附件,cad,三維,中英文,翻譯 is Rollover mitigation control Unified chassis control velopmen rabili ed perfo controller, and the human driver are investigated through a full-scale driving simulator on the VTT which consists of a real-time vehicle simulator, a visual animation engine, a visual display, and suitable by a small a disproportionately the vehicle control system. Accordingly, in 2002, NHTSA time- and method for rollover prevention that employs an optimal tire force ARTICLE IN PRESS Contents lists available at ScienceDirect Control Engineering Control Engineering Practice 18 (2010) 585–597 (Yoon, Kim, fax: +82 2 882 0561. automotive industry as it does not consider the effects of suspension deflection, tire traction aspects, or the dynamics of Liu proposed a robust active suspension for rollover prevention (Yang and (2) the type which indirectly influences roll motions by controlling the yaw Vehicle stability Virtual test track Design and evaluation of a unified chass prevention and vehicle stability improvement Jangyeol Yoon a , Wanki Cho a , Juyong Kang a , Bongyeong a School of Mechanical and Aerospace Engineering, Seoul National University, 599 Gwanangno, b Mando Corporation Central R it also calculates the desired braking force and the desired yaw moment for its objectives. Each control mode generates a control yaw moment and a longitudinal tire force in line with its coherent objective. The lower-level controller calculates the longitudinal and lateral tire forces as inputs of the control modules, such as the ESC and the AFS. 2.1. The upper-level controller: decision, desired braking force, and desired yaw moment The upper-level controller consists of three control modes and a switching logic. A control yaw moment and the longitudinal tire force are determined in line with its coherent control mode so that the switching across control modes is performed on the basis of the threshold. Based on the driver’s input and sensor signals, the upper-level controller determines which control mode is to be selected, as shown in Fig. 2. In this study, RI is used to detect an impending vehicle rollover where the RI is a dimensionless number that can indicate the risk of vehicle rollover and it is calculated through: the measured lateral acceleration, a y , the estimated roll angle, ^ f, the estimated roll rate, ^ _ f, and their critical values which depend on the vehicle geometry in the following manner (Yoon et al., 2007): In (1), C 1 , C 2 , and k 1 are positive constants (0oC 1 o1, 0oC 2 o1), C 1 and C 2 are weighting factors, which are related to the roll states and the lateral acceleration of the vehicle, and k 1 is a design parameter which is determined by the roll angle-rate phase plane analysis. These parameters in (1) are determined through a simulation study undertaken under various driving situations and tuned such that an RI of 1 indicates wheel-lift-off. A detailed description for the determination of the RI is provided in previous research (Yoon et al., 2007). The lateral acceleration can easily be measured from sensors that already exist on a vehicle equipped with an ESC system. However, additional sensors are needed to measure the roll angle and the roll rate, although it is difficult and costly to directly measure these (Schubert, Nichols, Fig. 1. RI/VS-based UCC strategy. RI?C 1 fetT C12 C12 C12 C12_ f th t _ fetT C12 C12 C12 C12 C12 C12f th f th _ f th 0 @ 1 A tC 2 a y C12 C12 C12 C12 a y,c C18C19 te1C0C 1 C0C 2 T fetT C12 C12 C12 C12 ???????????????????????????????????? fetTeT 2 t _ fetT C16C17 2 r 0 B B @ 1 C C A , f _ fC0k 1 f C16C17 40 RI?0, f _ fC0k 1 f C16C17 r0 8 > > > > > : e1T J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597 587 Fig. 2. Control modes for the proposed UCC system. ARTICLE IN PRESS 012345678 Time [sec] No control @43.2mph Control @45.6mph Roll angle 012345678 Time [sec] No control @43.2mph Contro l@45.6mph Lateral acceleration No control @43.2mph Control @45.6mph -15 -10 -5 0 5 10 15 -15 -10 -5 0 5 10 15 1 1.5 Roll angle [deg/sec] ay [m/s] J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597588 Wallner, Kong, in this the sliding surface and the sliding condition are defined as follows: s 1 ?gC0g des , 1 2 d dt s 1 2 ?s 1 _ s 1 rC0Z 1 s 1 jj e9T where Z 1 is a positive constant. The equivalent control input that would achieve _s 1 ?0 is calculated as follows: M z,eq ?C0I z 2eC0a ^ C f tb ^ C r T I z bC0 2ea 2 ^ C f tb 2 ^ C r T I z v x gt 2a ^ C f I z D f ! e10T Finally, the desired yaw moment for satisfying the sliding condition regardless of the model uncertainty is determined as follows: M z ?M z,eq C0K 2 sat gC0g des F 1 C18C19 e11T where F 1 is a control boundary, and the gain, K 2 , which satisfies the sliding condition, is calculated as follows: K 2 ?I z F yf I z C0abC0a 2 gtaD f C12 C12 C12 C12 t F yr I z bbC0b 2 g C12 C12 C12 C12 t _g des C12 C12 C12 C12 tZ 2 C26C27 e12T 2.1.2. Desired braking force for rollover prevention (the ROM mode) If the RI increases to a predefined RI threshold value, which can predict an impending rollover, the ROM control input should be applied to the vehicle in order to prevent rollover. Rollover prevention control can be achieved through vehicle speed control and the desired braking force is determined in this section to control the speed. In addition, the desired yaw moment, as determined in the previous section, is also applied to the vehicle to improve the maneuverability and the lateral stability. As mentioned previously, since vehicle rollovers occur at large lateral accelerations, the desired lateral acceleration should be defined and can be determined from the RI (cf. Eq. (1)) as follows: a y,des ? 1 C 2 RI tar C0C 1 fetT C12 C12 C12 C12_ f th t _ fetT C12 C12 C12 C12 C12 C12f th f th _ f th 0 @ 1 A C0e1C0C 1 C0C 2 T fetT C12 C12 C12 C12 ???????????????????????????????????? fetTeT 2 t _ fetT C16C17 2 r 0 B B @ 1 C C A 8 > : 9 > = > ; a y,c e13T In (2), the target RI value, RI tar , is set to 0.6. The desired vehicle speed for obtaining the desired lateral acceleration is calculated from the lateral vehicle dynamics as follows (Yoon et al., 2009): v x,des ? 1 g a y,des C0 a y,m C0v x g C0C1C8C9 e14T The desired braking force to yield the desired vehicle speed is calculated through a planar model, as shown in Fig. 7, and through the sliding mode control law. Fig. 7 shows a planar vehicle model including the desired braking force, DF x and the dynamic equation for the x-axis is described as follows: m _ v x ?F xr tF xf cosD f C0F yf sinD f tmv y gC0DF x e15T By the assumption of having small steering angles, Eq. (15) can be rewritten in terms of the derivative of the vehicle speed as follows: _ v x ? 1 m eF xr tF xf C0F yf D f Ttv y gC0 1 m DF x e16T Fig. 7. Planar model including the desired braking force. the use of braking because the ESC module has some negative ARTICLE IN PRESS effects as the simple distribution scheme determines only the differential braking input for the ESC module. These two schemes are switched in accordance with the protocol for switching across control modes in the upper-level controller and the only ESC module is used in the ROM mode since the optimized distribution scheme for the AFS and ESC modules provides a very small braking to each wheel, which cannot decrease the vehicle speed which is essential for preventing rollover. Moreover, the slip angle of the tire is proportionally increased with the lateral acceleration as shown in Fig. 8. Since vehicle rollovers generally occurs at large lateral acceleration, the slip angle of the tire is also very large in the ROM mode situation. The AFS module cannot generate the lateral tire force in large slip angle situations as shown in Fig. 9; therefore the AFS module is not used in the ROM mode, that is, the ESC is the most effective for the ROM mode. For this reason, only the ESC control module is used for the ROM mode. 2.2.1. Tire-force distribution in vehicle stability situations (ESC-g/ ESC-b mode) In vehicle stability situations that do not have risk of rollover, the control interventions for maneuverability, ESC-c, and for lateral stability, ESC-b, are activated. When the lateral accelera- tion is small enough so that the slip angle is small, the characteristics of the lateral tire force lie within the linear region, as shown in Fig. 9. In these situations, only the AFS control module is applied and the AFS control input is determined through the consideration of the 2-D bicycle model as follows: Slip angle [deg] ESCAFS + ESC AFS Lateral tire force [N] 0 36 Fig. 9. Characteristics of the lateral tire force. 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 -8 -6 -4 -2 0 Lateral acceleration [g] Slip angle [deg] Vehicle Stability Rollover Prevention Fig. 8. Relation between the lateral acceleration and the slip angle. J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597 591 Fig. 10. Coordinate system corresponding Dd f ? M z 2aC f e19T When the lateral acceleration increases greatly, the combined control inputs that are based on the ESC and AFS modules are applied. Since the ESC module has some negative effects, such as the degradation of ride comfort and the wear of tires and brakes, the optimized coordination of tire forces is focused on minimizing the use of braking. An optimal coordination of the lateral and longitudinal tire forces for the desired yaw moment is determined through the Karush–Kuhn–Tucker (KKT) conditions (Cho, Yoon, two of these constraints are determined as follows: fexT?C0 t 2 D 1 DF x1 taD 2 DF y1 C0M Z ?0 e23T gexT?eDF x1 tF x1 T 2 teDF y1 tF y1 T 2 C0m 2 F z1 2 r0 e24T In the above, D 1 ?1teF z3 =F z1 T, D 2 ?1teF z2 =F z1 T. The equality constraint in (23) means that the sum of the yaw moment generated by the longitudinal and the lateral tire forces should be equal to the desired yaw moment. The inequality constraint in (24) means that the sum of the long- itudinal and the lateral tire forces should be less than the friction forces on the tire. From (22)–(24), the Hamiltonian is defined as follows: H?DF x1 2 tl C0 t 2 D 1 DF x1 taD 2 DF y1 C0M Z C18C19 tr eDF x1 tF x1 T 2 teDF y1 tF y1 T 2 C0m 2 UF z1 2 tc 2 C16C17 e25T where l is the Lagrange multiplier, c the slack variable, and r the semi-positive number. First-order necessary conditions about the Hamiltonian are determined by the Karush–Kuhn–Tucker condition theory as follows: @H @DF x1 ?2DF x1 C0 t 2 D 1 lt2reDF x1 tF x1 T?0 e26T @H @DF y1 ?aD 2 lt2reDF y1 tF y1 T?0 e27T @H @l ?C0 t 2 D 1 DF x1 taD 2 DF y1 C0DM Z ?0 e28T rgexT?r eDF x1 tF x1 T 2 teDF y1 tF y1 T 2 C0m 2 F z1 2 C16C17 ?0 e29T J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597592 F xF,max xR,max F F xF F xR F zF μ μF zR Δ Fig. 11. Friction circles of the front and rear tires. Fig. 12. Hardware configuration of the driving From (29), two cases are derived with respect to r and g(x)as follows: Case 1. r?0, gexTo0. Case 2. r40, g(x)?0. Case 1 means that the sum of longitudinal and lateral tire forces is smaller than the friction of the tire. On the other hand, Case 2 means that the sum of the longitudinal and lateral tire forces is equal to the friction of the tire. The solutions of the optimization problem represented in (3.41) can be obtained for both cases. If the desire yaw moment is positive, M z 40, the solutions are obtained as follows: Case 1 : DF x1 ?0 DF y1 ? M Z aD 2 0 B @ e30T Case 2 : DF x1 ? C0eF x1 tkzTt ??????????????????????????????????????????????????????? e1tk 2 Tm 2 F z1 2 C0ekF x1 C0zT 2 q e1tk 2 T DF y1 ? tD 1 2aD 2 DF x1 t 1 aD 2 M Z e31T where k?(tD 1 /2aD 2 ) and z?(1/aD 2 )M Z +F y1 . The brake pressure for the ESC module and the additional steering angle for the AFS module are determined from (32) simulator with a human in-the-loop. ARTICLE IN PRESS as follows: DD f ? DF yi C f P Bi ? r wf DF xi K Bi ei?1,2T 0 B B B @ e32T In (32), K Bi is the brake gain, and r wf the radius of the wheel. When the desired yaw moment is negative, M z o0, the tire forces can be obtained in a manner similar to (30) and (31). 2.2.2. Tire-force distribution in rollover situations (ROM mode) In the previous sections, the desired braking force, which should be subjected to the vehicle for rollover prevention, and the desired yaw moment for reducing the error in the yaw rate have been determined. By utilizing the above two values, a braking- force distribution is accomplished simply to help prevent vehicle rollover, while ensuring that the vehicle follows the intended path of the driver. The forces of the vehicle can be determined kinematically, as follows: DF x,left ? 1 2 DF x t M z t 1 M z 8 > e33T 0 2 4 6 8 1012141618 Time [sec] Yaw rate 024681012141618 Time [sec] Lateral acceleration 0 2 4 6 8 10 12 14 16 18 -20 -10 0 10 20 -6 -4 -2 0 2 4 6 0 2 4 -50 0 50 Time [sec] Yaw rate [deg/s] Lateral acceleration [m/s 2 ] Steering wheel angle [deg] Vehicle test Simulator Vehicle test Simulator Vehicle test Simulator Steering wheel angle J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597 593 0 2 4 6 8 10 12 14 16 18 Time [sec] Vehicle test Simulator Roll angle -4 -2 Roll angle [deg] Fig. 13. Comparison between actual vehicle test data and the driving simulator (for the slalom test). DF x,right ? 2 DF x C0 t : The braking forces of the left and right sides are obtained by substituting (18) and (11) into (33). Fig. 11 shows the friction circles of the front and rear tires and the traction force, determined through the shaft torque, is applied at the front tire, and the drag force is applied at the rear tire. The maximum braking forces of the front and rear tires can be determined as follows: DF xf,max ?F xf C0 ?????????????????????????????? emF zf T 2 C0eF yf T 2 q e34T DF xr,max ?C0F xr C0 ?????????????????????????????? emF zr T 2 C0eF yr T 2 q e35T The braking-force distributions of the front and rear tires are achieved by using equations from (33) through to (35) as follows: DF xr,left ? DF xr,left,max C12 C12 C12 C12 DF xf,left,max C12 C12 C12 C12 DF xf,left e36T DF xr,right ? DF xr,right,max C12 C12 C12 C12 DF xf,right,max C12 C12 C12 C12 DF xf,right e37T In the above, DF xf,left tDF xr,left ?DF x,left and DF xf,right t DF xr,right ?DF x,right . 80km/h Obstacle Fig. 14. The test scenario: obstacle avoidance. ARTICLE IN PRESS The braking pressure of the front-left wheel can be determined as follows: P Bf,left ? r wf eDF xf,left T K Bf if DF xf,left oDF xf,max r wf eDF xf,max T K Bf if DF xf,left ZDF xf,max 8 > > > : e38T The other tire forces can be obtained in a manner similar to (38). 3. Full-scale driving simulator The configuration of the full-scale driving simulator for the human-in-the-loop system is shown in Fig. 12, consisting of four parts: a real-time (RT) simulation hardware, a visual graphical engine, a human-vehicle interface, and a motion platform. The host computer in Fig. 12 is utilized to modify the vehicle simulation program and to display the current vehicle status. The RT simulation hardware calculates the variables of the vehicle model represented using a CARSIM model controlled by the UCC controller with measured driver reactions. By the use of the vehicle-behavior information obtained using RT simulation hardware, the visual graphical engine projects a visual representation of the driving conditions to the human driver via a beam projector with a 100-in screen who interacts with the 3-D virtual simulation and the kinesthetic cues of the simulator body. The driver’s responses are acquired through the steering wheel angle, brake pressure, and throttle positioning sensors, as shown in Fig. 12. The motion platform provides kinesthetic cues, which are related to the behavior of the vehicle with regard to the human driver. An actual full-sized braking system, including a vacuum booster, master cylinder, calipers, etc., is implemented in the simulator so that the feel of the braking action is similar to that of an actual vehicular brake pedal. In the case of the steering wheel, a spring and damper are used to produce the reactive forces of the steering wheel where the spring and damper characteristics are adjusted to make the feel of the steering wheel similar to that of an actual vehicle being driven in the high-speed range. 3.1. Configurations of the driving simulator The most important feature of the driving simulator is to guarantee real-time performance and so all the subsystems are 8 0 -200 -100 100 200 100 120 10 Steering wheel angle [deg] w/o control RI-based ROM RI/VS-based UCC RI-based ROM 0 2 4 6 8 10 12 14 16 18 0 2 4 6 8 10 12 14 16 18 Time [sec] Time [sec] Steering wheel angle. Lateral acceleration. -1 -0.5 0 0.5 1 0 1.5 2 Lateral acceleration [g] w/o control RI-based ROM RI/VS-based UCC w/o control RI-based ROM RI/VS-based UCC J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597594 -10 -5 0 5 Roll angle [deg] RI/VS-based UCC 0 2 4 6 8 10 12 14 16 18 Time [sec] Roll angle. 0 2 4 6 8 101214161 0 20 40 60 80 Velocity [km/h] w/o control RI-based ROM RI/VS-based UCC w/o control Time [sec] Velocity. Fig. 15. Driving tests results using the full-scale 024681012141618 Time [sec] Rollover index. 0 2 4 6 8 10 12 14 16 18 -0.5 0 0.5 Time [sec] Yawrate error [deg/sec] w/o control RI-based ROM RI/VS-based UCC Yaw rate error. 0.5 1 Rollover index simulator based on the VTT. ARTICLE IN PRESS to the simulator body, as shown in Fig. 12 and the motion trajectories. If the UCC control input is not applied, the vehicle rolls over in this situation. It is clear from Fig. 15(e) that the RI increases over unity in the absence of control. Further, the roll angle and lateral acceleration also increase to large values, as shown in Fig. 15(c) and (d). In addition, because this situation is very severe, the vehicle deviates from the lane, as shown in Fig. 17. It can be seen that the driver’s detects the dropped obstacle at about five seconds and immediately tries to avoid the obstacle by changing lane. The vehicle velocities at about five seconds of three cases, viz., NON-control, RI-based ROM, and RI/VS-based UCC, are similar to each other, as shown in Fig. 15(b). When the UCC control is activated, two of the control systems yield good resistance to rollover, as shown in Fig. 15(c) and (e). As the RI-based ROM system intends to control the vehicle in a direction that is opposite to the driver’s intention, the yaw rate 0 2 4 6 8 1012141618 Time [sec] Brake pressures [MPa] Front-left Front-right Rear-left Rear-right RI-based ROM system. 0 2 4 6 8 1012141618 0 5 10 15 20 0 2 4 6 8 10 Time [sec] Brake pressures [MPa] Front-left Front-right Rear-left Rear-right RI/VS-based UCC system. Fig. 16. Brake pressures. J. Yoon et al. / Control Engineering Practice 18 (2010) 585–597 595 platform allows displacements up to a maximum of about 710 cm (heave) and 7101 (roll and pitch). The motion platform renders the linear and angular accelerations of the simulated vehicle model, as computed by the RT simulation hardware so that the human driver gets an impression that s/he is driving an actual vehicle by means of the kinesthetic cues generated by the motion platform, and from the visual representation of the driving situation provided by the visual graphical engine. 3.2. Validation of the vehicle simulator The driving simulator used in this paper is evaluated via actual vehicle test data and Fig. 13 shows the results of a slalom test in which the driver maintains an approximately constant vehicle speed of about 60 km/h. The cone width is 30 m. The magnitude and frequency of the driver’s steering inputs are almost identical in both the vehicle test results and the driving simulator, as shown in Fig. 13(a). The vehicle responses in terms of the yaw rate, the lateral acceleration and the roll angle are also quite similar to the actual test results as shown in Fig. 13(b)–(d). The comparison between the driving simulator and actual vehicle test results shows that the proposed driving simulator is feasi