裝配圖大學(xué)生方程式賽車設(shè)計（總體設(shè)計）（有cad圖+三維圖）

裝配圖大學(xué)生方程式賽車設(shè)計（總體設(shè)計）（有cad圖+三維圖）,裝配,大學(xué)生,方程式賽車,設(shè)計,總體,整體,cad,三維 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 f t C12 C12 C12 C12 f th f t C12 C12 C12 C12 C12 C12f th f th f th 0 1 A C 2 a y C12 C12 C12 C12 a y c C18C19 1C0C 1 C0C 2 f t C12 C12 C12 C12 f t 2 f t 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 1 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 9 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 2 C0a C f b C r I z bC0 2 a 2 C f b 2 C r I z v x g 2a C f I z D f 10 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 11 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 g aD f C12 C12 C12 C12 F yr I z bbC0b 2 g C12 C12 C12 C12 g des C12 C12 C12 C12 Z 2 C26C27 12 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 f t C12 C12 C12 C12 f th f t C12 C12 C12 C12 C12 C12f th f th f th 0 1 A C0 1C0C 1 C0C 2 f t C12 C12 C12 C12 f t 2 f t C16C17 2 r 0 B B 1 C C A 8 9 a y c 13 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 14 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 F xf cosD f C0F yf sinD f mv y gC0DF x 15 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 F xr F xf C0F yf D f v y gC0 1 m DF x 16 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 19 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 f x C0 t 2 D 1 DF x1 aD 2 DF y1 C0M Z 0 23 g x DF x1 F x1 2 DF y1 F y1 2 C0m 2 F z1 2 r0 24 In the above D 1 1 F z3 F z1 D 2 1 F z2 F z1 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 l C0 t 2 D 1 DF x1 aD 2 DF y1 C0M Z C18C19 r DF x1 F x1 2 DF y1 F y1 2 C0m 2 UF z1 2 c 2 C16C17 25 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 l 2r DF x1 F x1 0 26 H DF y1 aD 2 l 2r DF y1 F y1 0 27 H l C0 t 2 D 1 DF x1 aD 2 DF y1 C0DM Z 0 28 rg x r DF x1 F x1 2 DF y1 F y1 2 C0m 2 F z1 2 C16C17 0 29 J Yoon et al Control Engineering Practice 18 2010 585 597592 F xF max xR max F F xF F xR F zF F zR 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 g x o0 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 30 Case 2 DF x1 C0 F x1 kz 1 k 2 m 2 F z1 2 C0 kF x1 C0z 2 q 1 k 2 DF y1 tD 1 2aD 2 DF x1 1 aD 2 M Z 31 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 i 1 2 0 B B B 32 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 M z t 1 M z 8 33 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 mF zf 2 C0 F yf 2 q 34 DF xr max C0F xr C0 mF zr 2 C0 F yr 2 q 35 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 36 DF xr right DF xr right max C12 C12 C12 C12 DF xf right max C12 C12 C12 C12 DF xf right 37 In the above DF xf left DF xr left DF x left and DF xf right 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 DF xf left K Bf if DF xf left oDF xf max r wf DF xf max K Bf if DF xf left ZDF xf max 8 38 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 feasible for describing actual vehicle dynamic behaviors This means that the driving simulator accurately reproduces actual driving conditions 4 Evaluation of the proposed UCC based on a VTT Tests using the full scale driving simulator based on the VTT have been conducted to verify the proposed RI VS based UCC control algorithm and its performance with that of the previous RI based ROM control system are compared The tests based on the VTT have been conducted by thirteen drivers and the results are analyzed and summarized here The test scenario is set to the obstacle avoidance situation shown in Fig 14 so that when a driver follows the preceding vehicle moving at a constant speed of 90 km h in a straight lane and an object is dropped suddenly from the preceding vehicle In this situation the driver abruptly steers the vehicle to avoid the dropped obstacle and the vehicle is placed in a dangerous situation Moreover in this extreme situation vehicle rollover is possible and there may be a loss of maneuverability without an UCC control system Tests have been conducted by thirteen drivers Fig 1