CN107092245B - Hardware-in-loop simulation test bed for automobile dynamic chassis control system - Google Patents

Hardware-in-loop simulation test bed for automobile dynamic chassis control system Download PDF

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CN107092245B
CN107092245B CN201710402031.3A CN201710402031A CN107092245B CN 107092245 B CN107092245 B CN 107092245B CN 201710402031 A CN201710402031 A CN 201710402031A CN 107092245 B CN107092245 B CN 107092245B
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vehicle
wheels
wheel
module
modeling
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CN107092245A (en
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马畅
窦传威
孙安宁
魏宏
熊云亮
吴光强
张亮修
王宇
郭炯珉
秦洋洋
金杰
鞠丽娟
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SAIC Volkswagen Automotive Co Ltd
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Tongji University
SAIC Volkswagen Automotive Co Ltd
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    • G—PHYSICS
    • G05—CONTROLLING; REGULATING
    • G05B—CONTROL OR REGULATING SYSTEMS IN GENERAL; FUNCTIONAL ELEMENTS OF SUCH SYSTEMS; MONITORING OR TESTING ARRANGEMENTS FOR SUCH SYSTEMS OR ELEMENTS
    • G05B23/00—Testing or monitoring of control systems or parts thereof
    • G05B23/02—Electric testing or monitoring
    • G05B23/0205—Electric testing or monitoring by means of a monitoring system capable of detecting and responding to faults
    • G05B23/0218—Electric testing or monitoring by means of a monitoring system capable of detecting and responding to faults characterised by the fault detection method dealing with either existing or incipient faults
    • G05B23/0243—Electric testing or monitoring by means of a monitoring system capable of detecting and responding to faults characterised by the fault detection method dealing with either existing or incipient faults model based detection method, e.g. first-principles knowledge model
    • G—PHYSICS
    • G05—CONTROLLING; REGULATING
    • G05B—CONTROL OR REGULATING SYSTEMS IN GENERAL; FUNCTIONAL ELEMENTS OF SUCH SYSTEMS; MONITORING OR TESTING ARRANGEMENTS FOR SUCH SYSTEMS OR ELEMENTS
    • G05B2219/00—Program-control systems
    • G05B2219/20—Pc systems
    • G05B2219/24—Pc safety
    • G05B2219/24065—Real time diagnostics

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  • General Physics & Mathematics (AREA)
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Abstract

The invention discloses a hardware-in-loop simulation test bed for a dynamic chassis control system of an automobile.A man-car-road closed-loop digital simulation model is built on a host machine based on a Matlab/Simulink platform, is converted into an executable C code through an RTW compiling module and is downloaded into a CPU (central processing unit) of a target machine, a DCC (data communication controller) keeps communication with the target machine through an I/O (input/output) data conversion module, acquires man-car-road closed-loop digital model data in real time, outputs of the DCC control a damper electromagnetic valve, and a current acquisition module acquires a control current signal of the damper electromagnetic valve in real time and feeds the control current signal back to the target machine through the I/O data conversion module to form a closed-loop; and the simulation test bed evaluates the control effects under different working conditions and different modes, and gives a corresponding evaluation result after each simulation is finished. The invention has the advantage that the dynamic control of the automobile chassis is realized by adaptively adjusting the damping forces of the four shock absorbers.

Description

Hardware-in-loop simulation test bed for automobile dynamic chassis control system
Technical Field
The invention relates to a simulation test bed, in particular to a hardware-in-loop simulation test bed for an automobile dynamic chassis control system.
Background
The Dynamic Chassis Control system (DCC) is also called as an adaptive Chassis Control system, and can realize adaptive variable adjustment of four suspension dampers according to road conditions, driving conditions and requirements of a driver, and adjust an automobile Chassis into three modes, namely a standard mode, a Sport mode and a Comfort mode. The automobile provided with the DCC dynamic chassis control system can feel unprecedented driving comfort on the basis of keeping clear road feel, and the exercise chassis or the comfort chassis is correspondingly selected according to different driving environments, so that the chassis can perfectly match the driving conditions with the intention of a driver in real time all the time and maintain the balance of the driving conditions. DCC has compromise riding comfort and manipulation stability simultaneously through the design conflict of adjustable shock absorber and electronic power assisted steering solution motion chassis and travelling comfort chassis, can effectively solve car operating stability and ride comfort technical problem.
The public proposes a Dynamic Chassis Control (DCC) system, the system adopts MONROE (Chinese is translated into a trillione) valve Control continuous damping adjustable shock absorber under Tiannake flag, the controller is developed by German continent and the public, can realize the self-adaptive variable adjustment of four suspension damping according to the road surface condition, the driving condition and the requirements of drivers, adjusts the automobile Chassis into three modes of 'Normal type' (Normal), 'Sport' (sports) and 'Comfort' (Comfort), solves the design conflict of the Sport Chassis and the Comfort Chassis through the adjustable shock absorber and the electric power steering, and can effectively solve the technical problems of the operation stability and the Comfort riding of the automobile.
The university of fertilizer industry proposes an automobile chassis integrated control system and control method (200810021298.9). The control system detects wheel speed signals, torque signals, engine rotating speed signals, vertical acceleration signals, brake pedal signals and the like of an automobile through sensors, and inputs the signals into a main coordination CPU, the main coordination CPU respectively transmits the signals to an ABSCPU, an EPSCPU and an ASSCPU, and simultaneously sends out coordination commands according to the analysis of the signals, and the ABSCPU, the EPSCPU and the ASSCPU control corresponding driving modules according to the received sensor signals and the coordination commands. The invention overcomes the problem of mutual interference among the EPS, ASS and ABS systems in the prior automobile, realizes the coordination control of the three systems, and comprehensively improves the driving smoothness, safety and operation stability of the automobile. The university of Tongji proposed an automobile chassis integrated controller hardware-in-the-loop simulation test bench (200810040444.2) that integrated the functions of the anti-lock braking system (ABS), Traction Control System (TCS), and direct yaw moment control (DYC) for hardware-in-the-loop testing.
As a novel and practical technology, the dynamic chassis control system (DCC) realizes the dynamic control of the automobile chassis of the automobile by adaptively adjusting the damping forces of the four shock absorbers, and is greatly different from the prior invention in the aspects of an entire automobile modeling method, a control algorithm, an actuator and the like.
Disclosure of Invention
The invention aims to provide an xPC Target-based real-time platform, which realizes real-time communication between a shock absorber solenoid valve and a man-vehicle-road closed-loop digital simulation model and a DCC controller, and the running state of the shock absorber solenoid valve is controlled by the DCC controller.
In order to solve the technical problems, the invention provides a hardware-in-loop simulation test bed of an automobile dynamic chassis control system, which comprises a host machine, a target machine, a monitor, an I/O data conversion module, a network interface card, a USBCAN interface card, a BDM downloader, a DCC controller, a damper solenoid valve and a current sampling module, wherein the host machine is provided with a man-vehicle-road closed-loop digital simulation model based on a Matlab/Simulink platform, the man-vehicle-road closed-loop digital simulation model is converted into an executable C code through an RTW compiling module and downloaded into a CPU of the target machine, the DCC controller is communicated with the target machine through the I/O data conversion module, the DCC controller acquires man-vehicle-road closed-loop digital model data in the target machine in real time, the output of the DCC controller controls the damper solenoid valve, and the current acquisition module acquires a control current signal of the damper solenoid valve in real time, feeding back to the target machine through an I/O data conversion module to form a closed loop; and the simulation test bed evaluates the control effects under different working conditions and different modes, and gives a corresponding evaluation result after each simulation is finished.
The invention provides a vehicle longitudinal-lateral-vertical dynamics unified modeling idea, which is used for building a human-vehicle-road closed loop digital simulation model on a host machine based on a Matlab/Simulink platform, so that the built dynamic model is representative, and the mathematical theory analysis and simulation modeling of a vehicle longitudinal-lateral-vertical dynamics nonlinear model are realized on the basis of analyzing the behavior characteristics of complex nonlinear dynamics coupled by multiple vehicle systems, and the method comprises the following steps: 1) modeling assumptions; 2) modeling a power transmission system; 3) modeling a vehicle body; 4) modeling a suspension; 5) modeling a tire; 6) modeling of the driver.
1) Modeling assumptions:
generally, the higher the complexity of the model or the more degrees of freedom, the higher the simulation accuracy, but the numerical computation amount will increase and affect the real-time performance of the simulation. Therefore, taking into account the necessary vehicle dynamics coupling factors, it is necessary to make corresponding assumptions simplifications. Coupling factors that must be considered during vehicle motion are:
Figure 502205DEST_PATH_IMAGE001
the vehicle yaw motion caused by the steering of the wheels is coupled with each other kinematically and dynamically;
Figure 525394DEST_PATH_IMAGE002
the interaction between the tire and the road surface is not negligible, the distribution of its longitudinal and lateral tire forces being affected by the traction ellipses;
Figure 452898DEST_PATH_IMAGE003
the coupling exists between the longitudinal-lateral-vertical motion of the vehicle, the vertical load transfer of the vehicle can be caused by the longitudinal and lateral acceleration motion of the vehicle, so that the vertical dynamics of the vehicle is influenced, and the adhesion characteristic and the lateral deviation characteristic of the tire can be influenced by the change of the vertical load, so that the braking performance and the operation stability of the whole vehicle are influenced.
To simplify the modeling process, the following assumptions are made on the basis of fully considering vehicle coupling and strong non-linearity:
1. simplifying the modeling process of the power transmission system; 2. neglecting the influence of asymmetric wheel alignment parameters, and assuming that the center distance and the wheel distance of the suspension are equal; 3. assuming that the roll center and the pitch center are both located on the vehicle longitudinal bisecting plane and the roll axis is located above the pitch axis; 4. neglecting the roll and pitch motions of the unsprung mass; 5. it is assumed that the unsprung mass and the sprung mass are resiliently connected in the vertical direction and rigidly connected in the horizontal direction.
2) Modeling a power transmission system:
in order to comprehensively represent the unstable state process of the engine in the actual working process of the vehicle, a first-order inertia link with a hysteresis characteristic is added on the basis of the stable output characteristic of the engine to obtain the dynamic torque characteristic of the engine, namely:
Figure 404805DEST_PATH_IMAGE004
(1)
in the formula (I), the compound is shown in the specification,
Figure 50550DEST_PATH_IMAGE005
in order to output the torque dynamically from the engine,
Figure 877429DEST_PATH_IMAGE006
hair with indicationSteady state torque characteristic function of engine, which is engine speed
Figure 659441DEST_PATH_IMAGE007
And throttle opening degree
Figure 782249DEST_PATH_IMAGE008
Is a function of the non-linear function of (c),
Figure 915290DEST_PATH_IMAGE009
is a time constant, taken here
Figure 968696DEST_PATH_IMAGE010
。
The dynamic relation between the output torque and the output rotating speed of the engine is as follows:
Figure 182378DEST_PATH_IMAGE011
(2)
in the formula (I), the compound is shown in the specification,
Figure 476087DEST_PATH_IMAGE012
effective rotational inertia of the rotating part of the engine and the clutch part;
Figure 96424DEST_PATH_IMAGE013
is the engine rotational angular acceleration;
Figure 999527DEST_PATH_IMAGE014
outputting torque for an engine flywheel;
Figure 756130DEST_PATH_IMAGE015
to input torque to the clutch.
The vehicle under study was equipped with a dual clutch automatic transmission, and the engine output torque was considered equal to the transmission input torque, i.e., the input torque of the transmission, regardless of the engagement/disengagement process of the dual clutches during the modeling process
Figure 486320DEST_PATH_IMAGE016
(3)
In the formula (I), the compound is shown in the specification,
Figure 328374DEST_PATH_IMAGE017
effective rotational inertia of a rotating part of a transmission and a transmission shaft at a certain gear;
Figure 35168DEST_PATH_IMAGE018
and
Figure 911857DEST_PATH_IMAGE019
transmitting angular acceleration and angular velocity for a certain gear of the transmission;
Figure 547369DEST_PATH_IMAGE020
total drive torque for the wheels;
Figure 876719DEST_PATH_IMAGE021
is the transmission speed ratio;
Figure 387204DEST_PATH_IMAGE022
the speed ratio of the main speed reducer is obtained;
Figure 118399DEST_PATH_IMAGE023
for the transmission efficiency of the transmission system;
Figure 924812DEST_PATH_IMAGE024
is the wheel angular velocity.
Total drive torque
Figure 741459DEST_PATH_IMAGE025
Is simultaneously applied to two front wheels
Figure 790055DEST_PATH_IMAGE026
The wheel rotation dynamics equation is as follows:
Figure 375757DEST_PATH_IMAGE027
(4)
in the formula (I), the compound is shown in the specification,
Figure 618651DEST_PATH_IMAGE028
equivalent moment of inertia for the wheel;
Figure 922593DEST_PATH_IMAGE029
and
Figure 509301DEST_PATH_IMAGE030
the angular velocity and the angular acceleration of the wheel are respectively;
Figure 215089DEST_PATH_IMAGE031
is the tire longitudinal force;
Figure 363305DEST_PATH_IMAGE032
is the effective radius of the tire;
Figure 154543DEST_PATH_IMAGE033
and
Figure 544942DEST_PATH_IMAGE034
driving torque and braking torque of the wheels respectively;
Figure 105236DEST_PATH_IMAGE035
the wheel rotation damping coefficient;
Figure 689933DEST_PATH_IMAGE036
corresponding to the left front, right front, left rear and right rear wheels, respectively.
3) Vehicle body modeling
The vehicle body comprises a sprung mass and an unsprung mass, and the longitudinal-lateral-vertical unified dynamic model of the vehicle is established based on Lagrangian analytical mechanics.
Vehicle coordinate system
Figure 702888DEST_PATH_IMAGE037
Origin of (2)
Figure 982732DEST_PATH_IMAGE038
And center of pitch
Figure 335216DEST_PATH_IMAGE039
Center of coincidence, roll
Figure 90814DEST_PATH_IMAGE040
Relative to
Figure 591065DEST_PATH_IMAGE041
Satisfy the requirement of
Figure 323267DEST_PATH_IMAGE042
And (4) relationship. Sprung mass coordinate system
Figure 671203DEST_PATH_IMAGE043
Origin of (2)
Figure 846969DEST_PATH_IMAGE044
Coinciding with the sprung mass centre of mass, the unsprung mass mainly corresponds to four unsprung masses. Inertial frame
Figure 83784DEST_PATH_IMAGE045
Vehicle coordinate system
Figure 370409DEST_PATH_IMAGE046
And sprung mass coordinate system
Figure 244955DEST_PATH_IMAGE047
Can be converted into each other. If directional cosine matrix is used
Figure 591623DEST_PATH_IMAGE048
Representing the above-mentioned coordinate rotation transformation, i.e.
Figure 4150DEST_PATH_IMAGE049
(5)
The conversion relationship between the inertial, vehicle and sprung mass coordinate systems is:
Figure 343733DEST_PATH_IMAGE050
(6)
according to the foregoing definition and analysis, the body portion contains a total of 6 degrees of freedom, namely 3 degrees of freedom in the longitudinal, lateral and yaw shared by the unsprung mass and the sprung mass, and 3 degrees of freedom in the roll, pitch and vertical directions shared by the sprung mass. The translational and rotational angular velocities of the sprung and unsprung masses are separately determined and the respective kinetic and potential energies are then represented.
According to the coordinate conversion relation, the sprung mass center of mass (the origin of the sprung mass coordinate system)
Figure 587633DEST_PATH_IMAGE044
) Relative to the inertial coordinate system
Figure 590355DEST_PATH_IMAGE051
Absolute position vector of point
Figure 552495DEST_PATH_IMAGE052
And absolute velocity vector
Figure 492507DEST_PATH_IMAGE053
Respectively as follows:
Figure 607224DEST_PATH_IMAGE054
(7)
Figure 30115DEST_PATH_IMAGE055
(8)
in the formula (I), the compound is shown in the specification,
Figure 728819DEST_PATH_IMAGE056
is in an inertial coordinate system
Figure 160937DEST_PATH_IMAGE057
Dot relative to
Figure 864582DEST_PATH_IMAGE058
A position vector of the point;
Figure 723954DEST_PATH_IMAGE059
as in the vehicle coordinate system
Figure 644374DEST_PATH_IMAGE060
Dot relative to
Figure 880183DEST_PATH_IMAGE061
The position vector of the point, expressed as:
Figure 703914DEST_PATH_IMAGE062
(9)
in the formula (I), the compound is shown in the specification,
Figure 734187DEST_PATH_IMAGE063
as vectors
Figure 141903DEST_PATH_IMAGE064
A component of (a);
Figure 650245DEST_PATH_IMAGE065
is composed of
Figure 594061DEST_PATH_IMAGE066
Relative to each other
Figure 795236DEST_PATH_IMAGE067
Vertical distance;
Figure 690248DEST_PATH_IMAGE068
is composed of
Figure 2281DEST_PATH_IMAGE069
Relative to each other
Figure 535024DEST_PATH_IMAGE070
The vertical distance is set according to the distance,
Figure 172679DEST_PATH_IMAGE071
。
under the inertial coordinate system
Figure 554988DEST_PATH_IMAGE072
The translational velocity of the point, i.e.
Figure 405132DEST_PATH_IMAGE073
(10)
Noting the angular velocity of the sprung mass about its own reference axis as
Figure 323541DEST_PATH_IMAGE074
Then, then
Figure 600938DEST_PATH_IMAGE075
(11)
The kinetic energy of the sprung mass comprises both translational and rotational parts of the sprung mass, namely:
Figure 470543DEST_PATH_IMAGE076
(12)
in the formula (I), the compound is shown in the specification,
Figure 327641DEST_PATH_IMAGE077
is the sprung mass;
Figure 631714DEST_PATH_IMAGE078
for sprung mass about its centre of mass
Figure 611171DEST_PATH_IMAGE079
The inertia tensor, taking into account the sprung mass about
Figure 702493DEST_PATH_IMAGE080
Plane symmetry, then
Figure 160019DEST_PATH_IMAGE081
Comprises the following steps:
Figure 787441DEST_PATH_IMAGE082
(13)
in the formula (I), the compound is shown in the specification,
Figure 672220DEST_PATH_IMAGE083
is the center of mass of sprung mass
Figure 516417DEST_PATH_IMAGE084
The moment of inertia or the product of inertia.
Substituting the formulas (10), (11) and (13) for the formula (12) to obtain the sprung mass kinetic energy
Figure 512055DEST_PATH_IMAGE085
：
Figure 728404DEST_PATH_IMAGE086
(14)
Similarly, the unsprung mass kinetic energy is composed of the translation and rotation of the unsprung mass and the runout of four wheels, namely:
Figure 315243DEST_PATH_IMAGE087
(15)
the total kinetic energy being sprung mass kinetic energy
Figure 115578DEST_PATH_IMAGE088
And unsprung mass kinetic energy
Figure 914906DEST_PATH_IMAGE089
To sum, i.e.
Figure 313658DEST_PATH_IMAGE090
。
The potential energy of the vehicle body comprises gravitational potential energy generated by height change of the sprung mass
Figure 320666DEST_PATH_IMAGE091
Figure 93450DEST_PATH_IMAGE092
(16)
In the formula (I), the compound is shown in the specification,
Figure 447202DEST_PATH_IMAGE093
vertical displacement from sprung mass centre to unsprung mass centre;
Figure 152990DEST_PATH_IMAGE094
for sprung mass at its equilibrium point position
Figure 799740DEST_PATH_IMAGE093
The value of (c).
The total kinetic energy, potential energy and dissipation energy of the vehicle body are introduced into a Lagrange equation, and then partial derivatives of the total kinetic energy, potential energy and dissipation energy are solved, so that a motion equation of the vehicle body can be obtained, wherein the Lagrange equation of the vehicle body is as follows:
Figure 590979DEST_PATH_IMAGE095
(17)
in the formula (I), the compound is shown in the specification,
Figure 482843DEST_PATH_IMAGE096
is a generalized coordinate under an inertial coordinate system;
Figure 43137DEST_PATH_IMAGE097
is a generalized force in an inertial coordinate system.
In general, the motion of the vehicle is used to be described in the vehicle coordinate system, and the following relationship is used to convert the generalized variable in the formula (18) into the generalized variable in the vehicle coordinate system.
Figure 946544DEST_PATH_IMAGE098
(18)
In the formula (I), the compound is shown in the specification,
Figure 975811DEST_PATH_IMAGE099
generalized coordinates under a vehicle coordinate system;
Figure 920633DEST_PATH_IMAGE100
is a generalized force in a vehicle coordinate system.
So far, a kinetic equation of the six-degree-of-freedom vehicle body model is obtained
Figure 584701DEST_PATH_IMAGE101
(19)
In the formula (I), the compound is shown in the specification,
Figure 589566DEST_PATH_IMAGE102
，
Figure 840550DEST_PATH_IMAGE103
and
Figure 323484DEST_PATH_IMAGE104
in the form of a matrix of coefficients,
Figure 107638DEST_PATH_IMAGE105
generalized coordinates under a vehicle coordinate system;
Figure 17825DEST_PATH_IMAGE106
is a generalized force in a vehicle coordinate system.
If the air resistance is to be neglected,
Figure 818422DEST_PATH_IMAGE107
primarily from ground tire forces and suspension forces,
Figure 354315DEST_PATH_IMAGE108
expressed as:
Figure 743708DEST_PATH_IMAGE109
(20)
in the formula (I), the compound is shown in the specification,
Figure 28059DEST_PATH_IMAGE110
in the form of a matrix of coefficients,
Figure 253635DEST_PATH_IMAGE111
in the tire coordinate system for four wheels
Figure 78371DEST_PATH_IMAGE112
And
Figure 571538DEST_PATH_IMAGE113
directional tire forces, derived from a tire model;
Figure 902157DEST_PATH_IMAGE114
suspension forces for four wheels were obtained from the suspension model.
The motion of the vehicle under the inertial coordinate system is obtained through the following kinematic relationship:
Figure 864296DEST_PATH_IMAGE115
(21)
in the formula (I), the compound is shown in the specification,
Figure 741992DEST_PATH_IMAGE116
,
Figure 840398DEST_PATH_IMAGE117
for the whole vehicle
Figure 279600DEST_PATH_IMAGE118
Longitudinal, lateral speed of the shaft;
Figure 729036DEST_PATH_IMAGE119
,
Figure 144843DEST_PATH_IMAGE120
vehicle edge
Figure 363335DEST_PATH_IMAGE121
Longitudinal, lateral speed of the shaft;
Figure 973439DEST_PATH_IMAGE122
is the yaw angle of the vehicle.
4) Suspension modeling:
the purpose of the suspension model is here to find the suspension forces and the vertical loads of the wheels and to give the vertical equations of motion of the unsprung mass. The suspension force includes the elastic force of the elastic member, the damping force of the damping member, and the vertical acting force of the stabilizer bar, and the suspension force corresponding to each wheel is expressed as
Figure 644591DEST_PATH_IMAGE123
(22)
In the formula (I), the compound is shown in the specification,
Figure 129668DEST_PATH_IMAGE124
is the stiffness coefficient of the elastic element;
Figure 202666DEST_PATH_IMAGE125
is a damping force of the shock absorber, which is controlled by a control current
Figure 983672DEST_PATH_IMAGE126
Relative speed of motion of damper
Figure 453705DEST_PATH_IMAGE127
(ii) related;
Figure 962047DEST_PATH_IMAGE128
a vertical acting force generated for the transverse stabilizer bar;
Figure 905863DEST_PATH_IMAGE129
vertical displacement for four wheels;
Figure 107037DEST_PATH_IMAGE130
vertical displacement of sprung mass and four suspension contacts, from vehicle pitch angle
Figure 2050DEST_PATH_IMAGE131
Side inclination angle
Figure 314083DEST_PATH_IMAGE132
And calculating the geometrical parameters of the vehicle.
Damping force of the shock absorber
Figure 112405DEST_PATH_IMAGE133
The relationship between the control current and the relative movement speed of the damper is shown in FIGS. 4-5.
The contact force between the wheel and the ground is
Figure 484481DEST_PATH_IMAGE134
(23)
In the formula (I), the compound is shown in the specification,
Figure 866790DEST_PATH_IMAGE135
the contact force between the four wheels and the ground is the dynamic load of the wheels moving vertically;
Figure 982513DEST_PATH_IMAGE136
respectively the stiffness coefficient of each wheel,
Figure 635343DEST_PATH_IMAGE137
and inputting the road surface corresponding to the four wheels.
Under the action of suspension force and contact force between the wheel and the ground, the vertical motion equation of the unsprung mass is
Figure 489904DEST_PATH_IMAGE138
(24)
The vertical load of the wheel being constituted by the static normal force, the longitudinal load transfer, the lateral load transfer and the dynamic load of the tyre, i.e.
Figure 844662DEST_PATH_IMAGE139
(25)
In the formula (I), the compound is shown in the specification,
Figure 514809DEST_PATH_IMAGE140
vertical load for four wheels;
Figure 271412DEST_PATH_IMAGE141
the vertical loads of four wheels under the static state of the vehicle;
Figure 922973DEST_PATH_IMAGE142
and
Figure 279874DEST_PATH_IMAGE143
the vertical load variation of the wheels caused by the longitudinal load transfer and the lateral load transfer of the vehicle respectively;
Figure 737400DEST_PATH_IMAGE144
is the tire dynamic load of four wheels.
5) Tire modeling:
the tire model is a mathematical relationship description between six component forces of the tire and the wheel motion parameters. The invention uses MF tyre model to obtain the generalized force acting on the vehicle body in the form of
Figure 99243DEST_PATH_IMAGE145
(26)
Easily known, tire force
Figure 249601DEST_PATH_IMAGE146
Load perpendicular to wheel
Figure 828219DEST_PATH_IMAGE147
Longitudinal slip ratio
Figure 823857DEST_PATH_IMAGE148
Tire slip angle
Figure 305785DEST_PATH_IMAGE149
Road surface adhesion coefficient
Figure 627045DEST_PATH_IMAGE150
And camber angle of the wheel
Figure 427379DEST_PATH_IMAGE151
It is related.
6) Modeling a driver:
during simulation, the speed and the driving direction of the vehicle dynamic model need to be controlled so as to ensure that the speed and the driving track of the vehicle conform to expected values. The speed control being PID control, i.e.
Figure 492287DEST_PATH_IMAGE152
(27)
In the formula (I), the compound is shown in the specification,
Figure 828722DEST_PATH_IMAGE153
setting the vehicle speed;
Figure 55304DEST_PATH_IMAGE154
the actual vehicle speed;
Figure 608514DEST_PATH_IMAGE155
is a desired acceleration; control parameter
Figure 211534DEST_PATH_IMAGE156
，
Figure 402475DEST_PATH_IMAGE157
，
Figure 65537DEST_PATH_IMAGE158
。
The driving direction control of the vehicle dynamic model adopts an optimal curvature driver model, and the relationship between the characteristic parameters of the driver and the parameters of the vehicle model is established according to the operating characteristics of the driver.
The system comprises an I/O data conversion module, a DCC controller and a CAN bus, wherein the I/O data conversion module comprises an I/O data conversion card and a CAN conversion card, the I/O data conversion card converts various dynamic parameter signals of the vehicle obtained by calculation of a target computer from digital quantity to analog quantity, a vehicle height sensor signal and a vehicle vertical acceleration sensor signal are directly sent to the DCC controller, and the rest signals are packaged into CAN data by the CAN conversion card and sent to the network interface card and then transmitted to the DCC controller through the CAN bus; the I/O data conversion card simultaneously converts the analog quantity output by the current sampling module into digital quantity to be sent to the target machine to form a closed loop.
The monitoring machine monitors and collects data on the CAN bus in real time through the CAN conversion card, and performs post-processing and analysis on the data.
The DCC controller comprises an MC9S12XDP512 minimum system, a signal input module and an output driving module, wherein the MC9S12XDP512 minimum system comprises a power module, a clock circuit, a reset circuit and a BDM interface circuit, the signal input module comprises a filter circuit module, a voltage division circuit module and a CAN signal receiving and transmitting circuit module, and the output driving module comprises a PWM module, an electromagnetic valve driving circuit module and a current feedback circuit module; the input signals of the DCC controller comprise a vehicle height sensor signal, an acceleration sensor signal, a DCC mode selection signal and a CAN signal; in the simulation process of the DCC system, the change of the damping force of each shock absorber and the change of the control current of the shock absorber are given, the control strategy is verified in real time, and the control parameters are adjusted until a satisfactory control effect is obtained.
The electromagnetic valves of the shock absorber comprise four proportional electromagnetic valves, PWM and I/O ports output by a control chip are adopted for control, and the duty ratio of the PWM is changed to control the opening degree of a valve core of the proportional electromagnetic valves, so that the damping force output by the shock absorber is changed.
The current sampling module comprises a high-precision sampling resistor, a high-impedance amplifier and a filter circuit, the high-precision sampling resistor is connected in series in a driving circuit of the proportional solenoid valve, the high-impedance amplifier amplifies the voltage at two ends of the sampling resistor, the voltage is filtered by the filter circuit and then input into the I/O data conversion card, and the current working current of the proportional solenoid valve is fed back.
The network interface card is a multi-node CAN communication card, and CAN signal transmission from the CAN conversion card to the DCC controller and the USBCAN interface card is realized.
And the USBCAN interface card collects data on the CAN bus in real time and sends the data to the monitor.
The invention has the following advantages:
1) hardware of a controller and an actuator of the dynamic chassis control system is in a loop, and prediction results of various control strategies are more definite;
2) in the early development stage of a controller of a dynamic chassis control system, a hardware-in-the-loop simulation test bed is adopted, and various control parameters, particularly control parameters under extreme dangerous working conditions, can be optimized;
3) the smoothness of a vehicle with the assembled dynamic chassis control system, the anti-roll stability of the curve working condition, the pitching attitude control of the starting working condition, the tire adhesion characteristic and the transverse stability under the emergency working condition can be tested;
4) the test environment is simplified, and various performances obtained by testing and the obtained optimized parameters are closer to those of the real vehicle test;
5) the vehicle running state is simulated by the real-time processing platform running simulation model, comprehensive and systematic testing is carried out on the automobile dynamic chassis control system hardware, the number of real vehicle road test testing times is reduced, the risk of test faults is effectively reduced, the development time is shortened, and the cost is reduced.
Drawings
The accompanying drawings, which are incorporated in and constitute a part of this application, illustrate embodiments of the invention and, together with the description, serve to explain the invention and not to limit the invention. In the drawings:
FIG. 1 is a functional block diagram of the present invention;
FIG. 2 is a schematic block diagram of the host of the present invention;
FIG. 3 is a schematic view of the vehicle body kinematics analysis of the present invention;
FIG. 4 is a graphical representation of the damping characteristics of the front shock absorber of the present invention;
FIG. 5 is a graph illustrating the damping characteristics of the rear shock absorber of the present invention;
FIG. 6 is a schematic block diagram of the CAN converter card according to the present invention;
figure 7 is a schematic block diagram of a DCC controller circuit in accordance with the present invention;
FIG. 8 is a schematic block circuit diagram of the current sampling module of the present invention;
fig. 9 is a flow chart of the operation of the present invention.
Detailed Description
The embodiments of the invention will be described in detail below with reference to the drawings, but the invention can be implemented in many different ways as defined and covered by the claims.
Embodiments of the present invention will be described in detail below with reference to the accompanying drawings.
Fig. 1 shows a schematic block diagram of an embodiment of the invention. As shown in figure 1, the invention provides a hardware-in-the-loop simulation test bed for an automobile dynamic chassis control system, which comprises a host machine 1, a target machine 2, a monitor 3, an I/O data conversion card 4, a CAN conversion card 5, a network interface card 6, a USBCAN interface card 7, a BDM downloader 8, a DCC controller 9, a shock absorber solenoid valve 10 and a current sampling module 11. On a host machine 1, a man-car-road closed-loop digital simulation model is built based on a Matlab/Simulink platform, the man-car-road closed-loop digital simulation model is converted into executable C codes through an RTW compiling module and downloaded to a CPU of a target machine 2 through an Ethernet, a DCC controller 9 keeps communication with the target machine 2 through an I/O data conversion card 4, man-car-road closed-loop digital model information in the target machine is collected in real time, four shock absorber electromagnetic valves 10 are controlled, a current collecting module 9 collects control current signals of the four shock absorber electromagnetic valves in real time and feeds the control current signals back to the target machine 2 through the I/O data conversion card 4, and a closed-loop is formed. LabVIEW graphical data acquisition software is installed on the monitor 3, data on a CAN bus is monitored and acquired in real time through the CAN conversion card 5, and post-processing and analysis are carried out on the data. The software code inside the DCC controller may be written on host 1 or other PC and sintered into the DCC controller 9 through BDM 8.
Based on the software and hardware composition, the hardware-in-loop simulation test bed of the automobile dynamic chassis control system controlled by the DCC controller is established.
As shown in fig. 2, the host 1 is a PC installed with Matlab/Simulink and Visual C + + target language compiler software environment, and a man-car-road closed-loop digital simulation model is established on the host 1 and can be converted into an executable C code by an RTW compiling module.
In order to make the established dynamics model representative, the invention provides a unified modeling idea of vehicle longitudinal-lateral-vertical dynamics, which realizes mathematical theory analysis and simulation modeling of the vehicle longitudinal-lateral-vertical dynamics nonlinear model on the basis of analyzing the behavior characteristics of complex nonlinear dynamics coupled by multiple systems of a vehicle, and comprises the following steps: 1) modeling assumptions; 2) modeling a power transmission system; 3) modeling a vehicle body; 4) modeling a suspension; 5) modeling a tire; 6) modeling of the driver.
1) Modeling assumptions:
generally, the higher the complexity of the model or the more degrees of freedom, the higher the simulation accuracy, but the numerical computation amount will increase and affect the real-time performance of the simulation. Therefore, taking into account the necessary vehicle dynamics coupling factors, it is necessary to make corresponding assumptions simplifications. Coupling factors that must be considered during vehicle motion are:
1. the vehicle yaw motion caused by the steering of the wheels is coupled with each other kinematically and dynamically; 2. the interaction between the tire and the road surface is not negligible, the distribution of its longitudinal and lateral tire forces being affected by the traction ellipses; 3. the coupling exists between the longitudinal-lateral-vertical motion of the vehicle, the vertical load transfer of the vehicle can be caused by the longitudinal and lateral acceleration motion of the vehicle, so that the vertical dynamics of the vehicle is influenced, and the adhesion characteristic and the lateral deviation characteristic of the tire can be influenced by the change of the vertical load, so that the braking performance and the operation stability of the whole vehicle are influenced.
To simplify the modeling process, the following assumptions are made on the basis of fully considering vehicle coupling and strong non-linearity:
1. simplifying the modeling process of the power transmission system; 2. neglecting the influence of asymmetric wheel alignment parameters, and assuming that the center distance and the wheel distance of the suspension are equal; 3. assuming that the roll center and the pitch center are both located on the vehicle longitudinal bisecting plane and the roll axis is located above the pitch axis; 4. neglecting the roll and pitch motions of the unsprung mass; 5. it is assumed that the unsprung mass and the sprung mass are resiliently connected in the vertical direction and rigidly connected in the horizontal direction.
2) Modeling a power transmission system:
in order to comprehensively represent the unstable state process of the engine in the actual working process of the vehicle, a first-order inertia link with a hysteresis characteristic is added on the basis of the stable output characteristic of the engine to obtain the dynamic torque characteristic of the engine, namely:
Figure 863901DEST_PATH_IMAGE004
(1)
in the formula (I), the compound is shown in the specification,
Figure 83661DEST_PATH_IMAGE005
in order to output the torque dynamically from the engine,
Figure 643956DEST_PATH_IMAGE006
representing the steady-state torque characteristic function of the engine, which is the engine speed
Figure 461608DEST_PATH_IMAGE007
And throttle opening degree
Figure 677826DEST_PATH_IMAGE008
Is a function of the non-linear function of (c),
Figure 622648DEST_PATH_IMAGE009
is a time constant, taken here
Figure 788181DEST_PATH_IMAGE010
。
The dynamic relation between the output torque and the output rotating speed of the engine is as follows:
Figure 527467DEST_PATH_IMAGE011
(2)
in the formula (I), the compound is shown in the specification,
Figure 276986DEST_PATH_IMAGE012
effective rotational inertia of the rotating part of the engine and the clutch part;
Figure 25499DEST_PATH_IMAGE013
is the engine rotational angular acceleration;
Figure 45539DEST_PATH_IMAGE014
outputting torque for an engine flywheel;
Figure 221305DEST_PATH_IMAGE015
to input torque to the clutch.
The vehicle under study was equipped with a dual clutch automatic transmission, and the engine output torque was considered equal to the transmission input torque, i.e., the input torque of the transmission, regardless of the engagement/disengagement process of the dual clutches during the modeling process
Figure 458120DEST_PATH_IMAGE016
(3)
In the formula (I), the compound is shown in the specification,
Figure 744745DEST_PATH_IMAGE017
effective rotational inertia of a rotating part of a transmission and a transmission shaft at a certain gear;
Figure 884871DEST_PATH_IMAGE018
and
Figure 231538DEST_PATH_IMAGE019
transmitting angular acceleration and angular velocity for a certain gear of the transmission;
Figure 378486DEST_PATH_IMAGE020
total drive torque for the wheels;
Figure 780386DEST_PATH_IMAGE021
is the transmission speed ratio;
Figure 775018DEST_PATH_IMAGE022
the speed ratio of the main speed reducer is obtained;
Figure 292587DEST_PATH_IMAGE023
for the transmission efficiency of the transmission system;
Figure 503995DEST_PATH_IMAGE024
is the wheel angular velocity.
Total drive torque
Figure 132422DEST_PATH_IMAGE025
Is simultaneously applied to two front wheels
Figure 981560DEST_PATH_IMAGE026
The wheel rotation dynamics equation is as follows:
Figure 607714DEST_PATH_IMAGE027
(4)
in the formula (I), the compound is shown in the specification,
Figure 791570DEST_PATH_IMAGE028
equivalent moment of inertia for the wheel;
Figure 535273DEST_PATH_IMAGE029
and
Figure 504497DEST_PATH_IMAGE030
the angular velocity and the angular acceleration of the wheel are respectively;
Figure 363869DEST_PATH_IMAGE031
is the tire longitudinal force;
Figure 284289DEST_PATH_IMAGE032
is the effective radius of the tire;
Figure 254519DEST_PATH_IMAGE033
and
Figure 78250DEST_PATH_IMAGE034
driving torque and braking torque of the wheels respectively;
Figure 46206DEST_PATH_IMAGE035
the wheel rotation damping coefficient;
Figure 204655DEST_PATH_IMAGE036
corresponding to the left front, right front, left rear and right rear wheels, respectively.
3) Vehicle body modeling
The vehicle body comprises a sprung mass and an unsprung mass, and the longitudinal-lateral-vertical unified dynamic model of the vehicle is established based on Lagrangian analytical mechanics.
Vehicle coordinate system
Figure 227844DEST_PATH_IMAGE037
Origin of (2)
Figure 233977DEST_PATH_IMAGE038
And center of pitch
Figure 435151DEST_PATH_IMAGE039
Center of coincidence, roll
Figure 64584DEST_PATH_IMAGE040
Relative to
Figure 642196DEST_PATH_IMAGE041
Satisfy the requirement of
Figure 440519DEST_PATH_IMAGE042
And (4) relationship. Sprung mass coordinate system
Figure 812595DEST_PATH_IMAGE043
Origin of (2)
Figure 929324DEST_PATH_IMAGE044
Coinciding with the sprung mass centre of mass, the unsprung mass mainly corresponds to four unsprung masses. Inertial frame
Figure 45048DEST_PATH_IMAGE045
Vehicle coordinate system
Figure 697877DEST_PATH_IMAGE046
And sprung mass coordinate system
Figure 240854DEST_PATH_IMAGE047
Can be converted into each other. If directional cosine matrix is used
Figure 110458DEST_PATH_IMAGE048
Representing the above-mentioned coordinate rotation transformation, i.e.
Figure 764294DEST_PATH_IMAGE049
(5)
The conversion relationship between the inertial, vehicle and sprung mass coordinate systems is:
Figure 537209DEST_PATH_IMAGE050
(6)
according to the foregoing definition and analysis, the body portion contains a total of 6 degrees of freedom, namely 3 degrees of freedom in the longitudinal, lateral and yaw shared by the unsprung mass and the sprung mass, and 3 degrees of freedom in the roll, pitch and vertical directions shared by the sprung mass. The translational and rotational angular velocities of the sprung and unsprung masses are separately determined and the respective kinetic and potential energies are then represented.
According to the coordinate conversion relation, the sprung mass center of mass (the origin of the sprung mass coordinate system)
Figure 251087DEST_PATH_IMAGE044
) Relative to the inertial coordinate system
Figure 342408DEST_PATH_IMAGE051
Absolute position vector of point
Figure 799935DEST_PATH_IMAGE052
And absolute velocity vector
Figure 427356DEST_PATH_IMAGE053
Respectively as follows:
Figure 312136DEST_PATH_IMAGE054
(7)
Figure 890753DEST_PATH_IMAGE055
(8)
in the formula (I), the compound is shown in the specification,
Figure 151970DEST_PATH_IMAGE056
is in an inertial coordinate system
Figure 633898DEST_PATH_IMAGE057
Dot relative to
Figure 627262DEST_PATH_IMAGE058
A position vector of the point;
Figure 443909DEST_PATH_IMAGE059
as in the vehicle coordinate system
Figure 820401DEST_PATH_IMAGE060
Dot relative to
Figure 156835DEST_PATH_IMAGE061
The position vector of the point, expressed as:
Figure 383417DEST_PATH_IMAGE062
(9)
in the formula (I), the compound is shown in the specification,
Figure 936627DEST_PATH_IMAGE063
as vectors
Figure 211751DEST_PATH_IMAGE064
A component of (a);
Figure 730588DEST_PATH_IMAGE065
is composed of
Figure 331334DEST_PATH_IMAGE066
Relative to each other
Figure 856993DEST_PATH_IMAGE067
Vertical distance;
Figure 935807DEST_PATH_IMAGE068
is composed of
Figure 831124DEST_PATH_IMAGE069
Relative to each other
Figure 337191DEST_PATH_IMAGE070
The vertical distance is set according to the distance,
Figure 740360DEST_PATH_IMAGE071
。
under the inertial coordinate system
Figure 357286DEST_PATH_IMAGE072
The translational velocity of the point, i.e.
Figure 850715DEST_PATH_IMAGE073
(10)
Noting the angular velocity of the sprung mass about its own reference axis as
Figure 527684DEST_PATH_IMAGE074
Then, then
Figure 27936DEST_PATH_IMAGE075
(11)
The kinetic energy of the sprung mass comprises both translational and rotational parts of the sprung mass, namely:
Figure 353613DEST_PATH_IMAGE076
(12)
in the formula (I), the compound is shown in the specification,
Figure 295024DEST_PATH_IMAGE077
is the sprung mass;
Figure 549419DEST_PATH_IMAGE078
for sprung mass about its centre of mass
Figure 474649DEST_PATH_IMAGE079
The inertia tensor, taking into account the sprung mass about
Figure 433378DEST_PATH_IMAGE080
Plane symmetry, then
Figure 72039DEST_PATH_IMAGE081
Comprises the following steps:
Figure 231756DEST_PATH_IMAGE082
(13)
in the formula (I), the compound is shown in the specification,
Figure 378703DEST_PATH_IMAGE083
is the center of mass of sprung mass
Figure 469019DEST_PATH_IMAGE084
The moment of inertia or the product of inertia.
Substituting the formulas (10), (11) and (13) for the formula (12) to obtain the sprung mass kinetic energy
Figure 962186DEST_PATH_IMAGE085
：
Figure 479755DEST_PATH_IMAGE086
(14)
Similarly, the unsprung mass kinetic energy is composed of the translation and rotation of the unsprung mass and the runout of four wheels, namely:
Figure 113999DEST_PATH_IMAGE087
(15)
the total kinetic energy being sprung mass kinetic energy
Figure 821055DEST_PATH_IMAGE088
And unsprung mass kinetic energy
Figure 168729DEST_PATH_IMAGE089
To sum, i.e.
Figure 857199DEST_PATH_IMAGE090
。
The potential energy of the vehicle body comprises gravitational potential energy generated by height change of the sprung mass
Figure 791788DEST_PATH_IMAGE091
Figure 223906DEST_PATH_IMAGE092
(16)
In the formula (I), the compound is shown in the specification,
Figure 691666DEST_PATH_IMAGE093
vertical displacement from sprung mass centre to unsprung mass centre;
Figure 551037DEST_PATH_IMAGE094
for sprung mass at its equilibrium point position
Figure 972922DEST_PATH_IMAGE093
The value of (c).
The total kinetic energy, potential energy and dissipation energy of the vehicle body are introduced into a Lagrange equation, and then partial derivatives of the total kinetic energy, potential energy and dissipation energy are solved, so that a motion equation of the vehicle body can be obtained, wherein the Lagrange equation of the vehicle body is as follows:
Figure 943152DEST_PATH_IMAGE095
(17)
in the formula (I), the compound is shown in the specification,
Figure 265418DEST_PATH_IMAGE096
is a generalized coordinate under an inertial coordinate system;
Figure 295691DEST_PATH_IMAGE097
is a generalized force in an inertial coordinate system.
In general, the motion of the vehicle is used to be described in the vehicle coordinate system, and the following relationship is used to convert the generalized variable in the formula (18) into the generalized variable in the vehicle coordinate system.
Figure 267189DEST_PATH_IMAGE098
(18)
In the formula (I), the compound is shown in the specification,
Figure 978793DEST_PATH_IMAGE099
generalized coordinates under a vehicle coordinate system;
Figure 155566DEST_PATH_IMAGE100
is a generalized force in a vehicle coordinate system.
So far, a kinetic equation of the six-degree-of-freedom vehicle body model is obtained
Figure 356740DEST_PATH_IMAGE101
(19)
In the formula (I), the compound is shown in the specification,
Figure 815534DEST_PATH_IMAGE102
，
Figure 330829DEST_PATH_IMAGE103
and
Figure 362108DEST_PATH_IMAGE104
in the form of a matrix of coefficients,
Figure 734183DEST_PATH_IMAGE105
generalized coordinates under a vehicle coordinate system;
Figure 617957DEST_PATH_IMAGE106
is a generalized force in a vehicle coordinate system.
If the air resistance is to be neglected,
Figure 733680DEST_PATH_IMAGE107
primarily from ground tire forces and suspension forces,
Figure 885045DEST_PATH_IMAGE108
expressed as:
Figure 428022DEST_PATH_IMAGE109
(20)
in the formula (I), the compound is shown in the specification,
Figure 799091DEST_PATH_IMAGE110
in the form of a matrix of coefficients,
Figure 764511DEST_PATH_IMAGE111
in the tire coordinate system for four wheels
Figure 521115DEST_PATH_IMAGE112
And
Figure 251304DEST_PATH_IMAGE113
directional tire forces, derived from a tire model;
Figure 93358DEST_PATH_IMAGE114
suspension forces for four wheels were obtained from the suspension model.
The motion of the vehicle under the inertial coordinate system is obtained through the following kinematic relationship:
Figure 800152DEST_PATH_IMAGE115
(21)
in the formula (I), the compound is shown in the specification,
Figure 676841DEST_PATH_IMAGE116
,
Figure 312353DEST_PATH_IMAGE117
for the whole vehicle
Figure 641703DEST_PATH_IMAGE118
Longitudinal, lateral speed of the shaft;
Figure 152188DEST_PATH_IMAGE119
,
Figure 883383DEST_PATH_IMAGE120
vehicle edge
Figure 689797DEST_PATH_IMAGE121
Longitudinal, lateral speed of the shaft;
Figure 506443DEST_PATH_IMAGE122
is the yaw angle of the vehicle.
4) Suspension modeling:
the purpose of the suspension model is here to find the suspension forces and the vertical loads of the wheels and to give the vertical equations of motion of the unsprung mass. The suspension force includes the elastic force of the elastic member, the damping force of the damping member, and the vertical acting force of the stabilizer bar, and the suspension force corresponding to each wheel is expressed as
Figure 555039DEST_PATH_IMAGE123
(22)
In the formula (I), the compound is shown in the specification,
Figure 140741DEST_PATH_IMAGE124
is the stiffness coefficient of the elastic element;
Figure 383635DEST_PATH_IMAGE125
is a damping force of the shock absorber, which is controlled by a control current
Figure 687577DEST_PATH_IMAGE126
Relative speed of motion of damper
Figure 274285DEST_PATH_IMAGE127
(ii) related;
Figure 980073DEST_PATH_IMAGE128
a vertical acting force generated for the transverse stabilizer bar;
Figure 128289DEST_PATH_IMAGE129
vertical displacement for four wheels;
Figure 919527DEST_PATH_IMAGE130
vertical displacement of sprung mass and four suspension contacts, from vehicle pitch angle
Figure 145864DEST_PATH_IMAGE131
Side inclination angle
Figure 706158DEST_PATH_IMAGE132
And calculating the geometrical parameters of the vehicle.
Damping force of the shock absorber
Figure 290854DEST_PATH_IMAGE133
The relationship between the control current and the relative movement speed of the damper is shown in FIGS. 4-5.
The contact force between the wheel and the ground is
Figure 303810DEST_PATH_IMAGE134
(23)
In the formula (I), the compound is shown in the specification,
Figure 497900DEST_PATH_IMAGE135
the contact force between the four wheels and the ground is the dynamic load of the wheels moving vertically;
Figure 725750DEST_PATH_IMAGE136
respectively the stiffness coefficient of each wheel,
Figure 730615DEST_PATH_IMAGE137
and inputting the road surface corresponding to the four wheels.
Under the action of suspension force and contact force between the wheel and the ground, the vertical motion equation of the unsprung mass is
Figure 480134DEST_PATH_IMAGE138
(24)
The vertical load of the wheel being constituted by the static normal force, the longitudinal load transfer, the lateral load transfer and the dynamic load of the tyre, i.e.
Figure 900751DEST_PATH_IMAGE139
(25)
In the formula (I), the compound is shown in the specification,
Figure 435637DEST_PATH_IMAGE140
vertical load for four wheels;
Figure 362136DEST_PATH_IMAGE141
the vertical loads of four wheels under the static state of the vehicle;
Figure 349684DEST_PATH_IMAGE142
and
Figure 885576DEST_PATH_IMAGE143
the vertical load variation of the wheels caused by the longitudinal load transfer and the lateral load transfer of the vehicle respectively;
Figure 822439DEST_PATH_IMAGE144
is the tire dynamic load of four wheels.
5) Tire modeling:
the tire model is a mathematical relationship description between six component forces of the tire and the wheel motion parameters. The invention uses MF tyre model to obtain the generalized force acting on the vehicle body in the form of
Figure 169107DEST_PATH_IMAGE145
(26)
Easily known, tire force
Figure 893218DEST_PATH_IMAGE146
Load perpendicular to wheel
Figure 983534DEST_PATH_IMAGE147
Longitudinal slip ratio
Figure 978166DEST_PATH_IMAGE148
Tire slip angle
Figure 167839DEST_PATH_IMAGE149
Road surface adhesion coefficient
Figure 129979DEST_PATH_IMAGE150
And camber angle of the wheel
Figure 69991DEST_PATH_IMAGE151
It is related.
6) Modeling a driver:
during simulation, the speed and the driving direction of the vehicle dynamic model need to be controlled so as to ensure that the speed and the driving track of the vehicle conform to expected values. The speed control being PID control, i.e.
Figure 184708DEST_PATH_IMAGE152
(27)
In the formula (I), the compound is shown in the specification,
Figure 607599DEST_PATH_IMAGE153
setting the vehicle speed;
Figure 306303DEST_PATH_IMAGE154
the actual vehicle speed;
Figure 738421DEST_PATH_IMAGE155
is a desired acceleration; control parameter
Figure 442066DEST_PATH_IMAGE156
，
Figure 301438DEST_PATH_IMAGE157
，
Figure 221858DEST_PATH_IMAGE158
。
The driving direction control of the vehicle dynamic model adopts an optimal curvature driver model, and the relationship between the characteristic parameters of the driver and the parameters of the vehicle model is established according to the operating characteristics of the driver.
The target machine 2 is a 610H industrial personal computer and realizes the communication between the target machine 2 and the DCC controller 9 through a data conversion module.
The data conversion module comprises an I/O data conversion card 4 (PCL-818L and PCL-726) and a CAN conversion card 5. The I/O data conversion card 4 converts various dynamic parameter signals of the vehicle calculated by the target computer 2 from digital quantity to analog quantity, wherein the vehicle height sensor signal and the vehicle vertical acceleration sensor signal are directly received by the DCC controller 9, and the rest signals are packaged into CAN messages by the CAN conversion card 5 and sent to the network interface card 6 and then transmitted to the DCC controller 9 through the CAN bus. The I/O data conversion card 4 simultaneously converts the analog quantity output by the current sampling module 11 into digital quantity for receiving by the target machine 2, thereby forming a closed loop.
The circuit principle of the CAN conversion card 5 is shown in figure 6, according to the input requirement of a signal acquisition module of a DCC controller 9, the CAN conversion card is designed by taking a Freescale Feishakall 8-bit control chip as a core, and various dynamic parameter signals of a vehicle output by the I/O data conversion card 4 are converted into CAN messages to be sent to a network interface card 6 for the DCC controller 9 and a USBCAN interface card 7 to receive.
The circuit principle of the DCC controller 9 is shown in FIG. 7, and the invention develops and designs the DCC controller by itself with a Freescale 16-bit control chip MC9S12XDP512 as a core according to the characteristics of a DCC system, wherein input signals of the DCC controller comprise a vehicle body height sensor signal, an acceleration sensor signal, a DCC mode selection signal and a CAN signal. The DCC controller includes an MC9S12XDP512 minimum system, a signal input module, and an output driver module. The MC9S12XDP512 minimum system comprises a power module, a clock circuit, a reset circuit, a BDM interface circuit and the like; the signal input module comprises a filter circuit module, a voltage division circuit module and a CAN signal transceiving circuit module; the output driving module comprises a PWM module, an electromagnetic valve driving circuit module and a current feedback circuit module.
The damper solenoid valve 10 includes four proportional solenoid valves, and is controlled by PWM and I/O ports output by a control chip. The valve core opening degree of the proportional solenoid valve can be realized by using BTS5090 of Infineon (England flying) as a driving chip and controlling through an I/O port and changing the duty ratio of PWM, so that the damping force output by the shock absorber is changed.
The current sampling module 11 is shown in fig. 8 and includes a high-precision sampling resistor, a high-impedance amplifier, and a filter circuit. A high-precision sampling resistor is connected in series in a proportional solenoid valve driving circuit, voltages at two ends of the sampling resistor are amplified by using a high-impedance differential amplifier, and high-frequency noise in signals is reduced through an RC (resistor-capacitor) filter circuit. And finally, inputting the filtered signal into an I/O data conversion board card 4 of the host machine 1, so that the current working current of the proportional solenoid valve can be determined.
The network interface card 6 is a multi-node CAN communication card to realize CAN signal transmission from the CAN conversion card 5 to the DCC controller 9 and the USBCAN interface card 7.
The USBCAN interface card 7 is a ZLG USBCAN-II intelligent CAN interface card and is used for collecting messages on a CAN bus in real time.
The monitoring machine 3 is a PC machine provided with LabVIEW graphical data acquisition software, is connected with the network interface card 6 through the USBCAN interface card 7, acquires the interactive information of the target machine 2 and the DCC controller 9 in real time, monitors abnormal data in the test process, and stores the data so as to carry out post-processing and analysis.
The BDM8 is used for sintering control codes written on the host machine 1 or other PC machines into the DCC controller 9, so as to realize the read-write and erase operations of the microprocessor Flash, facilitate the online tracking and debugging of the operation of the control codes and improve the development efficiency of the controller.
Through the steps, a hardware-in-loop simulation test bed of the dynamic chassis control system is established, and the hardware-in-loop simulation test bed can run and evaluate the control parameters of the electric control unit. The man-vehicle-road closed-loop system model runs in the target machine 2, the DCC controller 9 controls the working state of the electromagnetic valve 10 according to vehicle information given by the target machine 2 in real time, such as height sensor signals, acceleration sensor signals, DCC mode selection signals, CAN signals and the like, the circuit acquisition module 9 feeds back the current of the corresponding shock absorber to the CPU of the target machine 2 through the data board card, and the monitoring machine 3 judges test results in real time through the USBCAN interface card 7.
As shown in fig. 9, which is a working flow chart of the present invention, the hardware-in-loop simulation test bed can evaluate the control effect under different working conditions and different modes, and each simulation can give a corresponding result for evaluation. In the simulation process of the DCC system, the change of the damping force of each shock absorber, the change of the control current of the shock absorber and the like can be comprehensively given, so that the control strategy is verified in real time, and the control parameters are adjusted until a satisfactory control effect is obtained.
In addition, the hardware-in-the-loop simulation test bed can also realize the optimized matching of parameters of each component of a vehicle chassis, a tire and a transmission system, can realize the debugging of control parameters of the vehicle under the extreme dangerous working condition, and can detect and debug the circuit fault of the designed electronic control unit 3.
Due to the fact that hardware of the DCC controller 9 and the shock absorber electromagnetic valve is in a ring, various tested performances and obtained optimized parameters are close to those of a real vehicle test, the times of the real vehicle test are obviously reduced, the development period is shortened, and meanwhile a large amount of development cost is saved.
The above description is only a preferred embodiment of the present invention, and is not intended to limit the present invention, and various modifications and changes may be made by those skilled in the art. Any modification, equivalent replacement, or improvement made within the spirit and principle of the present invention should be included in the protection scope of the present invention.

Claims (8)

1. The utility model provides an automobile dynamic chassis control system hardware is at ring simulation test bench which characterized in that: the simulation test bed comprises a host machine, a target machine, a monitor, an I/O data conversion module, a network interface card, a USBCAN interface card, a BDM downloader, a DCC controller, a shock absorber electromagnetic valve and a current sampling module, wherein a man-vehicle-road closed-loop digital simulation model is built on the host machine based on a Matlab/Simulink platform, is converted into an executable C code through an RTW compiling module and is downloaded into a CPU (central processing unit) of the target machine, the DCC controller is communicated with the target machine through the I/O data conversion module, acquires man-vehicle-road closed-loop digital model data in the target machine in real time, the output of the DCC controller controls the shock absorber electromagnetic valve, and the current acquisition module acquires a control current signal of the shock absorber electromagnetic valve in real time and feeds the control current signal back to the target machine through the I/O data conversion module to form a closed-; the simulation test bed evaluates the control effects under different working conditions and different modes, and gives corresponding evaluation results after each simulation is finished;
a man-vehicle-road closed loop digital simulation model is built on the host machine based on a Matlab/Simulink platform, and mathematical theory analysis and simulation modeling of a vehicle longitudinal-lateral-vertical dynamics nonlinear model are realized on the basis of analyzing the multi-system coupled complex nonlinear dynamics behavior characteristics of the vehicle, and the method comprises the following steps: 1) modeling assumptions; 2) modeling a power transmission system; 3) modeling a vehicle body; 4) modeling a suspension; 5) modeling a tire; 6) modeling a driver;
wherein:
1) modeling assumptions, including a) simplifying the powertrain modeling process; b) neglecting the influence of asymmetric wheel alignment parameters, and assuming that the center distance and the wheel distance of the suspension are equal; c) assuming that the roll center and the pitch center are both located on the vehicle longitudinal bisecting plane and the roll axis is located above the pitch axis; d) neglecting the roll and pitch motions of the unsprung mass; e) the unsprung mass and the sprung mass are assumed to be elastically connected in the vertical direction and rigidly connected in the horizontal direction;
2) modeling a power transmission system:
in order to comprehensively represent the unstable state process of the engine in the actual working process of the vehicle, a first-order inertia link with a hysteresis characteristic is added on the basis of the stable output characteristic of the engine to obtain the dynamic torque characteristic of the engine, namely:
Figure DEST_PATH_IMAGE001
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE002
in order to output the torque dynamically from the engine,
Figure DEST_PATH_IMAGE003
representing the steady-state torque characteristic function of the engine, which is the engine speed
Figure DEST_PATH_IMAGE004
And throttle opening degree
Figure DEST_PATH_IMAGE005
Is a function of the non-linear function of (c),
Figure DEST_PATH_IMAGE006
is a time constant, taken here
Figure DEST_PATH_IMAGE007
；
The dynamic relation between the output torque and the output rotating speed of the engine is as follows:
Figure DEST_PATH_IMAGE009
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE010
effective rotational inertia of the rotating part of the engine and the clutch part;
Figure DEST_PATH_IMAGE012
is the engine rotational angular acceleration;
Figure DEST_PATH_IMAGE013
outputting torque for an engine flywheel;
Figure DEST_PATH_IMAGE014
inputting torque for the clutch;
the vehicle under study was equipped with a dual clutch automatic transmission, and the engine output torque was considered equal to the transmission input torque, i.e., the input torque of the transmission, regardless of the engagement/disengagement process of the dual clutches during the modeling process
Figure DEST_PATH_IMAGE016
；
In the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE017
effective rotational inertia of a rotating part of a transmission and a transmission shaft at a certain gear;
Figure DEST_PATH_IMAGE019
and
Figure DEST_PATH_IMAGE020
transmitting angular acceleration and angular velocity for a certain gear of the transmission;
Figure DEST_PATH_IMAGE021
total drive torque for the wheels;
Figure DEST_PATH_IMAGE022
is the transmission speed ratio;
Figure DEST_PATH_IMAGE023
the speed ratio of the main speed reducer is obtained;
Figure DEST_PATH_IMAGE024
for the transmission efficiency of the transmission system;
Figure DEST_PATH_IMAGE025
is the wheel angular velocity;
total drive torque
Figure DEST_PATH_IMAGE026
Is simultaneously applied to two front wheels
Figure DEST_PATH_IMAGE027
The wheel rotation dynamics equation is as follows:
Figure DEST_PATH_IMAGE029
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE030
equivalent moment of inertia for the wheel;
Figure DEST_PATH_IMAGE031
and
Figure DEST_PATH_IMAGE033
the angular velocity and the angular acceleration of the wheel are respectively;
Figure DEST_PATH_IMAGE034
is the tire longitudinal force;
Figure DEST_PATH_IMAGE035
is the effective radius of the tire;
Figure DEST_PATH_IMAGE036
and
Figure DEST_PATH_IMAGE037
driving torque and braking torque of the wheels respectively;
Figure DEST_PATH_IMAGE038
the wheel rotation damping coefficient;
Figure DEST_PATH_IMAGE039
respectively corresponding to the left front wheel, the right front wheel, the left rear wheel and the right rear wheel;
3) vehicle body modeling
The vehicle body comprises a sprung mass and an unsprung mass, and a longitudinal-lateral-vertical unified dynamic model of the vehicle is established based on Lagrangian analytical mechanics; the body portion contains 6 degrees of freedom in total, namely 3 degrees of freedom in the longitudinal, lateral and yaw shared by the unsprung and sprung masses, and 3 degrees of freedom in the roll, pitch and vertical directions shared by the sprung mass; respectively solving the translational motion and the rotation angular velocity of the sprung mass and the unsprung mass, and then expressing the respective kinetic energy and potential energy;
the kinetic energy of the sprung mass comprising both translational and rotational parts of the sprung mass, i.e. the kinetic energy of the sprung mass
Figure DEST_PATH_IMAGE040
：
Figure DEST_PATH_IMAGE042
；
Similarly, the unsprung mass kinetic energy is composed of the translation and rotation of the unsprung mass and the runout of four wheels, namely:
Figure DEST_PATH_IMAGE044
；
the total kinetic energy being sprung mass kinetic energy
Figure DEST_PATH_IMAGE045
And unsprung mass kinetic energy
Figure DEST_PATH_IMAGE046
To sum, i.e.
Figure DEST_PATH_IMAGE047
；
The potential energy of the vehicle body comprises gravitational potential energy generated by height change of the sprung mass
Figure DEST_PATH_IMAGE048
：
Figure DEST_PATH_IMAGE049
；
In the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE050
vertical displacement from sprung mass centre to unsprung mass centre;
Figure DEST_PATH_IMAGE051
for sprung mass at its equilibrium point position
Figure DEST_PATH_IMAGE052
A value of (d);
the total kinetic energy, potential energy and dissipation energy of the vehicle body are introduced into a Lagrange equation, and then partial derivatives of the Lagrange equation are calculated to obtain a motion equation of the vehicle body, so that a kinetic equation of a six-degree-of-freedom vehicle body model is obtained:
Figure DEST_PATH_IMAGE054
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE055
，
Figure DEST_PATH_IMAGE056
and
Figure DEST_PATH_IMAGE057
in the form of a matrix of coefficients,
Figure DEST_PATH_IMAGE059
generalized coordinates under a vehicle coordinate system;
Figure DEST_PATH_IMAGE060
is the generalized force under the vehicle coordinate system;
if the air resistance is to be neglected,
Figure DEST_PATH_IMAGE061
primarily from ground tire forces and suspension forces,
Figure DEST_PATH_IMAGE062
expressed as:
Figure DEST_PATH_IMAGE063
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE064
in the form of a matrix of coefficients,
Figure DEST_PATH_IMAGE066
in the tire coordinate system for four wheels
Figure DEST_PATH_IMAGE067
And
Figure DEST_PATH_IMAGE068
directional tire forces, derived from a tire model;
Figure DEST_PATH_IMAGE069
the suspension forces corresponding to the four wheels are obtained by a suspension model;
the motion of the vehicle under the inertial coordinate system is obtained through the following kinematic relationship:
Figure DEST_PATH_IMAGE071
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE072
,
Figure DEST_PATH_IMAGE073
for the whole vehicle
Figure DEST_PATH_IMAGE074
Longitudinal, lateral speed of the shaft;
Figure DEST_PATH_IMAGE076
for the whole vehicle
Figure DEST_PATH_IMAGE077
Longitudinal, lateral speed of the shaft;
Figure DEST_PATH_IMAGE078
a yaw angle of the vehicle;
4) suspension modeling:
solving the suspension force and the vertical load of the wheel, and giving a vertical motion equation of the unsprung mass; the suspension force includes an elastic force of the elastic member, a damping force of the damping member, and a vertical acting force of the stabilizer bar, and the suspension force corresponding to each wheel is expressed as:
Figure DEST_PATH_IMAGE079
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE080
is the stiffness coefficient of the elastic element;
Figure DEST_PATH_IMAGE081
is a damping force of the shock absorber, which is controlled by a control current
Figure DEST_PATH_IMAGE082
Relative speed of motion of damper
Figure DEST_PATH_IMAGE083
(ii) related;
Figure DEST_PATH_IMAGE084
a vertical acting force generated for the transverse stabilizer bar;
Figure DEST_PATH_IMAGE085
vertical displacement for four wheels;
Figure DEST_PATH_IMAGE086
vertical displacement of sprung mass and four suspension contacts, from vehicle pitch angle
Figure DEST_PATH_IMAGE087
Side inclination angle
Figure DEST_PATH_IMAGE088
And calculating the geometric parameters of the vehicle;
the contact force between the wheels and the ground is as follows:
Figure DEST_PATH_IMAGE090
；
in the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE091
the contact force between the four wheels and the ground is the dynamic load of the wheels moving vertically;
Figure DEST_PATH_IMAGE092
respectively the stiffness coefficient of each wheel,
Figure DEST_PATH_IMAGE093
inputting the road surface corresponding to the four wheels;
under the action of suspension force and contact force between the wheels and the ground, the vertical motion equation of the unsprung mass is as follows:
Figure DEST_PATH_IMAGE095
；
the vertical load of the wheel being constituted by the static normal force, the longitudinal load transfer, the lateral load transfer and the dynamic load of the tyre, i.e.
Figure DEST_PATH_IMAGE096
；
In the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE097
vertical load for four wheels;
Figure DEST_PATH_IMAGE098
the vertical loads of four wheels under the static state of the vehicle;
Figure DEST_PATH_IMAGE099
and
Figure DEST_PATH_IMAGE100
the vertical load variation of the wheels caused by the longitudinal load transfer and the lateral load transfer of the vehicle respectively;
Figure DEST_PATH_IMAGE101
tire dynamic load for four wheels;
5) tire modeling:
the tire model is a mathematical relationship description between six component forces of the tire and the wheel motion parameters in the form of:
Figure DEST_PATH_IMAGE102
；
wherein the tyre force
Figure DEST_PATH_IMAGE103
Load perpendicular to wheel
Figure DEST_PATH_IMAGE104
Longitudinal slip ratio
Figure DEST_PATH_IMAGE105
Tire slip angle
Figure DEST_PATH_IMAGE106
Road surface adhesion coefficient
Figure DEST_PATH_IMAGE107
And camber angle of the wheel
Figure DEST_PATH_IMAGE108
(ii) related;
6) modeling a driver:
the speed and the driving direction of a vehicle dynamic model need to be controlled during simulation, and the speed control adopts PID control, namely
Figure DEST_PATH_IMAGE110
；
In the formula (I), the compound is shown in the specification,
Figure DEST_PATH_IMAGE111
setting the vehicle speed;
Figure DEST_PATH_IMAGE112
the actual vehicle speed;
Figure DEST_PATH_IMAGE113
is a desired acceleration; control parameter
Figure DEST_PATH_IMAGE114
，
Figure DEST_PATH_IMAGE115
，
Figure DEST_PATH_IMAGE116
；
The driving direction control of the vehicle dynamic model adopts an optimal curvature driver model, and the relationship between the characteristic parameters of the driver and the parameters of the vehicle model is established according to the operating characteristics of the driver;
the DCC controller gives the change of the damping force of each shock absorber and the change of the control current of the shock absorber in the simulation process of the DCC system, verifies the control strategy in real time and adjusts the control parameters until a satisfactory control effect is obtained.
2. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the system comprises an I/O data conversion module, a DCC controller and a CAN bus, wherein the I/O data conversion module comprises an I/O data conversion card and a CAN conversion card, the I/O data conversion card converts various dynamic parameter signals of the vehicle obtained by calculation of a target computer from digital quantity to analog quantity, a vehicle height sensor signal and a vehicle vertical acceleration sensor signal are directly sent to the DCC controller, and the rest signals are packaged into CAN data by the CAN conversion card and sent to the network interface card and then transmitted to the DCC controller through the CAN bus; the I/O data conversion card simultaneously converts the analog quantity output by the current sampling module into digital quantity to be sent to the target machine to form a closed loop.
3. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the monitoring machine monitors and collects data on the CAN bus in real time through the CAN conversion card, and performs post-processing and analysis on the data.
4. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the DCC controller comprises an MC9S12XDP512 minimum system, a signal input module and an output driving module, wherein the MC9S12XDP512 minimum system comprises a power module, a clock circuit, a reset circuit and a BDM interface circuit, the signal input module comprises a filter circuit module, a voltage division circuit module and a CAN signal receiving and transmitting circuit module, and the output driving module comprises a PWM module, an electromagnetic valve driving circuit module and a current feedback circuit module; the input signals of the DCC controller comprise a vehicle height sensor signal, an acceleration sensor signal, a DCC mode selection signal and a CAN signal.
5. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the electromagnetic valves of the shock absorber comprise four proportional electromagnetic valves, PWM and I/O ports output by a control chip are adopted for control, and the duty ratio of the PWM is changed to control the opening degree of a valve core of the proportional electromagnetic valves, so that the damping force output by the shock absorber is changed.
6. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the current sampling module comprises a high-precision sampling resistor, a high-impedance amplifier and a filter circuit, the high-precision sampling resistor is connected in series in a driving circuit of the proportional solenoid valve, the high-impedance amplifier amplifies the voltage at two ends of the sampling resistor, the voltage is filtered by the filter circuit and then input into the I/O data conversion card, and the current working current of the proportional solenoid valve is fed back.
7. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: the network interface card is a multi-node CAN communication card, and CAN signal transmission from the CAN conversion card to the DCC controller and the USBCAN interface card is realized.
8. The automobile dynamic chassis control system hardware-in-the-loop simulation test bed according to claim 1, characterized in that: and the USBCAN interface card collects data on the CAN bus in real time and sends the data to the monitor.
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