JP2010247633A - Damping force control device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a damping force control device capable of controlling damping force to integrally suppress vertical vibration, roll vibration and pitch vibration of a sprung member and to suppress calculation load. <P>SOLUTION: When a turning state of a vehicle is in an understeering state, damping force of a suspension device mounted at each wheel position of the sprung member HA is controlled by requested damping force calculated by using a rear wheel approximate three-wheel model. When the turning state of the vehicle is in an oversteering state, the damping force of the suspension device mounted at each wheel position of the sprung member HA is controlled by requested damping force calculated by using a front wheel approximate three-wheel model. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a damping force control device that controls a damping force generated by a suspension device of a vehicle.

  A vehicle suspension apparatus includes a spring and a damper (shock absorber) interposed between a sprung member and an unsprung member. The damper has a function of attenuating vibration between the sprung member and the unsprung member. The damping force control device controls the damping force generated by the suspension device by variably controlling the damping force characteristic (damping coefficient) of the damper.

  Skyhook control theory is well known as a control theory for controlling the damping force generated by the suspension device. According to the Skyhook control theory, when the product of the sprung-road relative speed and the sprung speed is positive, the damping coefficient of the damper is proportional to the ratio between the sprung-road relative speed and the sprung speed. The damping coefficient of the damper is controlled to be the lowest damping coefficient when the product is negative. When this skyhook control theory is applied to damping force control, the damping force rapidly decreases when a request to vibrate vibration (active request) is made by switching the product from positive to negative. This causes a large damping force fluctuation.

As another control theory of the damping force of the suspension device, a nonlinear _H∞ control theory is also known. According to the nonlinear H _∞ control theory, with respect to the damping force control system represented by state space representation that is designed to be a bilinear system based on the equation of motion obtained from the vehicle model, approximately satisfies the Riccati inequality Thus, the control law (feedback controller) is calculated. By controlling the damping force based on the control law, the vibration of the sprung member is suppressed and controlled. When the nonlinear _H∞ control theory is applied, the damping force (required damping force) changes smoothly. For this reason, compared with the case where the skyhook control theory is applied, a rapid fluctuation of the damping force can be suppressed.

In Patent Document 1, the nonlinear H _∞ control theory is applied to a damping force control system designed on the basis of a single-wheel model of a vehicle, whereby the damping forces of a plurality of (for example, four) suspension devices attached to the vehicle are individually determined. Discloses a damping force control device that performs independent control. The damping force control device described in Patent Literature 1 is a damping force control system (generalized) expressed by a state space expression designed based on the vertical motion of an unsprung member and an unsprung member derived from a single-wheel model of a vehicle. The damping force of each suspension device is controlled by applying nonlinear _H∞ control theory to the plant.

Patent Document 2 discloses a damping force that integrally controls the damping force of a suspension device provided at each wheel position by applying nonlinear _H∞ control to a damping force control system designed based on a four-wheel model of a vehicle. A control device is disclosed. The damping force control device described in Patent Document 2 is an up-and-down (heave) motion, a roll (around the longitudinal axis) motion, a pitch (around the left-right axis) motion of a sprung member derived from a four-wheel model of a vehicle, By applying nonlinear _H∞ control theory to a damping force control system (generalized plant) represented by a state space expression designed based on the vertical motion of the unsprung member, three directions (vertical direction, The damping force of each suspension device is integratedly controlled so as to suppress vibration in the roll direction and the pitch direction).

JP 2000-148208 A

JP 2006-160185 A

According to the invention described in Patent Document 2, vibration in the vertical direction, roll direction, and pitch direction of the sprung member is integrated by controlling the damping force based on the nonlinear _H∞ control theory using a four-wheel model. It is suppressed. As a result, the riding comfort at the control target position is improved. However, there is a problem that the calculation load increases because the degree of the matrix used for calculating the control law is large. One method for suppressing the calculation load is that each suspension device is considered as a one-degree-of-freedom vibration system in which only the sprung member is relatively displaced with respect to the road surface. This is a method for controlling damping force by applying nonlinear _H∞ control to a damping force control system represented by a state space expression designed based on pitch motion. According to this control method, the calculation load is expected to be reduced because the movement of the unsprung member is not taken into consideration. However, since a control system designed based on such a model is uncontrollable, a control law cannot be calculated.

Table 1 compares the controllability of each damping force control system, the amount of operation (control input) in the state space representation of the system, the number of state amounts, and the degree of freedom of the model that is the basis of the state space representation. is there.

  That the control system is controllable can be paraphrased as that there is an operation amount (control input) that can control the state quantity to a desired value with respect to the disturbance acting on the system. In other words, if the degree of freedom of the model that is the basis of the control system is not less than the degree of freedom of the sprung member due to disturbance input (when it is large or equal), it can be said that the control system is controllable. . Conversely, when the degree of freedom of the model is less than the degree of freedom of the sprung member due to disturbance input, the control system is non-controllable.

  As shown in Table 1 (1), the vehicle model is a single-wheel model, the degree of freedom of the sprung member is 1 degree of freedom, and the degree of freedom of the suspension device is 2 degrees of freedom (the vertical motion of the sprung member and the spring In the case of the vertical movement of the lower member, the control system represented by the model is controllable. The disturbance of this system is represented by the vertical displacement (road surface displacement) of the ground contact surface of the wheel. When this disturbance is converted into a displacement of the sprung member (mode conversion), the disturbance is represented by the vertical displacement of the sprung member. Therefore, the degree of freedom of the sprung member due to disturbance is one degree of freedom (vertical movement). On the other hand, the degree of freedom of the model is two degrees of freedom obtained by adding the degree of freedom of the unsprung member to the degree of freedom of the sprung member. Thus, the control system is controllable because the degree of freedom of the model is greater than the degree of freedom of the sprung member due to disturbance input.

  As shown in Table 1 (2), the vehicle model is a four-wheel model, the degree of freedom of the sprung member is 3 degrees of freedom (vertical motion, roll motion, pitch motion), and the degree of freedom of the suspension device is 2 degrees of freedom. In the case of degrees (sprung motion, unsprung motion), the control system represented by the model is controllable. The disturbance of this system is represented by the road surface displacement at each wheel (four wheel) position. When this disturbance is mode-converted into displacement of the sprung member, the disturbance is represented by vertical displacement, roll displacement, pitch displacement, and warp (twist) displacement of the sprung member. Therefore, the degree of freedom of the sprung member due to disturbance is originally 4 degrees of freedom. On the other hand, the degree of freedom of the model is 7 degrees of freedom by adding the degree of freedom of the unsprung member (vertical movement) at each wheel position to the degree of freedom of the sprung member (vertical movement, roll movement, pitch movement). is there. Thus, the control system is controllable because the degree of freedom of the model is greater than the degree of freedom of the sprung member due to disturbance input.

  As shown in (3) of Table 1, the vehicle model is a four-wheel model, the degree of freedom of the sprung member is 3 degrees of freedom (vertical motion, roll motion, pitch motion), and the degree of freedom of the suspension device is 1 freedom. In the case of the degree (up and down movement of the sprung member), the control system represented by the model is uncontrollable. The degree of freedom of the sprung member due to disturbance input of this system is 4 degrees of freedom as described above. On the other hand, the degree of freedom of the model is 3 degrees of freedom expressed by the motion of the sprung member (vertical motion, roll motion, pitch motion) (when the sprung member is assumed to be a rigid body in the model, the warp motion equation) Cannot be derived). Therefore, since the degree of freedom of the model (3 degrees of freedom) is less than the degree of freedom of the sprung member due to disturbance (4 degrees of freedom), this control system is uncontrollable.

As described above, when the damping force is controlled by applying the nonlinear H _∞ control theory to the four-wheel model, when the suspension device is represented as a one-degree-of-freedom vibration system model in order to suppress the calculation load, the model is represented by the model. Since the control system is disabled, the damping force cannot be controlled.

  The present invention has been made to address the above-described problems, and in a damping force control device that controls the damping force of a vehicle suspension device, the vertical vibration, roll vibration, and pitch vibration of the sprung member are integrally suppressed. In addition, an object of the present invention is to provide a damping force control device capable of controlling the damping force so that the calculation load is suppressed.

The present invention is characterized in that in the damping force control device for controlling the damping force of the four suspension devices attached to each wheel position of the sprung member of the vehicle, the two attached to the left and right front wheel positions of the sprung member. On a spring derived from a three-wheel model representing the motion of a vehicle in which two rear wheel suspension devices attached to the left and right rear wheel positions of the front wheel side suspension device or sprung member are replaced with one virtual suspension device. By applying the nonlinear _H∞ control theory to a control system designed based on the vertical motion, roll motion and pitch motion of the members, control targets to be generated by the three suspension devices represented in the three-wheel model The requested damping force, which is the damping force of the virtual suspension device, is calculated from the calculated requested damping force. The required damping force that calculates the required damping force for the four suspension devices attached to each wheel position of the sprung member by distributing to the required damping force for the two suspension devices replaced with the suspension device Calculating means, and damping force control means for controlling the damping force of the four suspension devices attached to each wheel position of the sprung member based on the requested damping force calculated by the requested damping force calculating means. The damping force control device is provided.

  In this case, the required damping force calculating means is calculated by a model required damping force calculating means for calculating required damping forces for the three suspension devices represented in the three-wheel model, and the model required damping force calculating means. The required damping force distribution means for distributing the required damping force for the virtual suspension device among the required damping forces to the required damping force for the two suspension devices replaced with the virtual suspension device. .

  According to the above invention, the required damping force for the four suspension devices attached to each wheel position of the sprung member is calculated based on the three-wheel model of the vehicle. In the three-wheel model, two virtual rear suspension devices attached to the left and right front wheel positions of the sprung member or two rear wheel suspension devices attached to the left and right rear wheel positions of the sprung member are combined into one virtual suspension. This is a vehicle model replaced with a device. When the vehicle model is a three-wheel model, the disturbance of the control system represented by the three-wheel model is the displacement (or displacement speed) of the ground contact surface of the wheels connected to the three suspension devices. When this disturbance is mode-converted into the displacement of the sprung member, the disturbance can be expressed by vertical displacement, roll displacement, and pitch displacement of the sprung member. Therefore, the degree of freedom of the sprung member by disturbance input is 3 degrees of freedom (vertical motion, roll motion, pitch motion). In the three-wheel model, the sprung member considered as a rigid body moves in the vertical direction, the roll direction, and the pitch direction. When each suspension device is a one-degree-of-freedom vibration system that considers only the motion of the sprung member, the degree of freedom of the model and the degree of freedom of the sprung member due to disturbance input are equal. For this reason, the control system represented by the motion (vertical motion, roll motion, pitch motion) derived from the three-wheel model is controllable.

When the vehicle model is a three-wheel model in this way, the control system represented by the model is controllable. Therefore, by applying the non-linear _H∞ control theory to a controllable system, the required damping force for the three suspension devices represented in the model can be calculated. Then, by distributing the required damping force for the virtual suspension device to the required damping forces for the two suspension devices replaced by the virtual suspension device, the four attached to each wheel position of the sprung member The required damping force for the suspension device can be calculated. Based on the required damping force for the four suspension devices calculated in this way, the damping force of each suspension device is controlled.

  Further, by using the three-wheel model, the controllability can be given to the system only by considering the movement of the sprung member in three directions. Therefore, the required damping force can be calculated without considering each suspension device as a one-degree-of-freedom vibration system and considering the equation of motion of the unsprung member. Therefore, it is possible to reduce the calculation load when calculating the required damping force.

  The three-wheel model may be a rear-wheel approximate three-wheel model representing a vehicle motion in which the two rear-wheel suspension devices are replaced with one rear-wheel virtual suspension device. Alternatively, the three-wheel model may be a front-wheel approximate three-wheel model representing a motion of a vehicle in which the two front wheel side suspension devices are replaced with one front wheel side virtual suspension device.

  Whether the rear wheel approximate three-wheel model or the front wheel approximate three-wheel model is used as the three-wheel model can be determined based on various vehicle characteristics. In general, when calculating the required damping force of the suspension device at each wheel position using a three-wheel model, the suspension device on the side that is replaced by the virtual suspension device (approximate side) is replaced with the suspension device that is not approximated. In comparison, the effect of suppressing roll vibration is low. Therefore, it is preferable to use a three-wheel model in which the suspension device on the side where the effect of suppressing the roll vibration is not regarded as important is replaced with a virtual suspension device. For example, the suspension device attached to the non-drive wheel side may be replaced with a virtual suspension device. That is, it is preferable to use a front wheel approximate three-wheel model for an FR vehicle and a rear wheel approximate three-wheel model for an FF vehicle. Further, the model may be determined in consideration of the priority of the ride comfort. For example, when it is desired to effectively suppress the roll vibration on the front seat side of the vehicle, the rear wheel approximate three-wheel model is used. When the roll vibration on the rear seat side is effectively suppressed, the front wheel approximate three-wheel model is used. Is good. Further, a three-wheel model may be used in which the suspension device on the side having a small amount of the sprung acceleration sensor that detects the vertical acceleration (sprung acceleration) of the sprung member is replaced with a virtual suspension device. For example, when two sprung acceleration sensors are attached to the front wheel side and one rear wheel side of the sprung member, a rear wheel approximate three-wheel model may be used.

Another feature of the present invention is that the required damping force calculation means is a rear wheel approximation 3 representing a vehicle motion in which the two rear wheel suspension devices are replaced with one rear wheel virtual suspension device. The non-linear _H∞ control theory is applied to a control system designed based on the vertical motion, roll motion and pitch motion of the sprung member derived from the wheel model. Calculate the required damping force for each suspension device, and distribute the required damping force for the rear wheel side virtual suspension device among the calculated required damping forces to the required damping force for the two rear wheel side suspension devices. A first required damping force calculating means for calculating the required damping force for the four suspension devices attached to each wheel position of the sprung member; Based on the vertical motion, roll motion and pitch motion of the sprung member derived from the front wheel approximate three-wheel model representing the motion of the vehicle in which the two front wheel suspension devices are replaced by one front wheel virtual suspension device. By applying nonlinear _H∞ control theory to the control system to be designed, the required damping force for the three suspension devices represented in the front wheel approximate three-wheel model is calculated, and the front wheel of the calculated required damping force is calculated. The required damping force for the four suspension devices attached to each wheel position of the sprung member is distributed by distributing the required damping force for the side virtual suspension device to the required damping force for the two front wheel suspension devices. A second required damping force calculating means for calculating a force, wherein the damping force control means is the first required damping force calculating means. The damping force of the four suspension devices attached to each wheel position of the sprung member is controlled based on the requested damping force calculated by any one of the second requested damping force calculation means.

  According to the above invention, the damping force control means is one of the required damping force calculated based on the rear wheel approximate three-wheel model and the required damping force calculated based on the front wheel approximate three-wheel model. Is used to control the damping force of the four suspension devices attached to the sprung member. For this reason, damping force control using the rear wheel approximate three-wheel model and damping force control using the front wheel approximate three-wheel model can be selected according to the running state of the vehicle, the turning state, and the user's preference.

  In this case, the damping force control means is configured to calculate the required damping force calculated by the first required damping force calculation means and the required damping force calculated by the second required damping force calculation means based on the turning state of the vehicle. Requested damping force selecting means for selecting one of them, and based on the requested damping force selected by the requested damping force selecting means, the damping forces of the four suspension devices attached to each wheel position of the sprung member It is good to control.

  Further, the required damping force selection means selects the required damping force calculated by the first required damping force calculation means when the turning state of the vehicle is an understeer state, and the turning state of the vehicle is an oversteer state. Sometimes it is preferable to select the required damping force calculated by the second required damping force calculating means.

  The damping force control device is detected by a yaw rate detection sensor that detects an actual yaw rate that occurs when the vehicle turns, target yaw rate calculation means that calculates a target yaw rate based on a vehicle speed and a steering angle, and the actual yaw rate detection sensor. A turning state determining means for determining whether the turning state of the vehicle is an understeer state or an oversteer state based on a yaw rate deviation representing a difference between the actual yaw rate and the target yaw rate calculated by the target yaw rate calculating means; The required damping force selection means may select the required damping force based on the turning state of the vehicle determined by the turning state determination means.

  When the turning state of the vehicle is an understeer state, the understeer state can be corrected by suppressing roll vibration on the front wheel side. In addition, when the damping force of the suspension device attached to each wheel position is controlled by applying the rear wheel approximate three-wheel model, the front wheel side suspension device attached to the left and right front wheel positions effectively controls the vehicle roll vibration. A damping force that suppresses the noise is generated. Therefore, the understeer state is corrected by applying the rear wheel approximate three-wheel model and controlling the damping force of the suspension device when the turning state of the vehicle is the understeer state. This improves the stability when turning the vehicle.

  On the other hand, when the turning state of the vehicle is the oversteer state, the oversteer state can be corrected by suppressing the roll vibration on the rear wheel side. In addition, when the damping force of the suspension device attached to each wheel position is controlled by applying the front wheel approximate three-wheel model, the rear wheel side suspension device attached to the left and right rear wheel positions has the effect of rolling the vehicle. A damping force that suppresses the noise is generated. Therefore, when the turning state of the vehicle is an oversteer state, the oversteer state is corrected by applying the front wheel approximate three-wheel model to control the damping force of the suspension device. This improves the stability when turning the vehicle.

  According to still another aspect of the present invention, the damping force control device includes a right front wheel-side sprung acceleration sensor that detects vertical acceleration of the sprung member at the right front wheel position, and a vertical acceleration of the sprung member at the left front wheel position. Left front wheel side sprung acceleration sensor, right rear wheel side sprung acceleration sensor for detecting the vertical acceleration of the sprung member at the right rear wheel position, and vertical acceleration of the sprung member at the left rear wheel position A left rear wheel-side sprung acceleration sensor, and the damping force control means is configured such that when either one of the right front wheel-side sprung acceleration sensor or the left front wheel-side sprung acceleration sensor is abnormal, Based on the required damping force calculated by the required damping force calculating means, the damping force of the suspension device attached to each wheel position of the sprung member is controlled, and the right rear wheel side sprung acceleration sensor When either one of the left rear wheel side sprung acceleration sensor is abnormal, the suspension attached to each wheel position of the sprung member based on the required damping force calculated by the first required damping force calculation means The purpose is to control the damping force of the device.

  According to the above invention, when the sprung acceleration sensor is attached to the front / rear and left / right wheel positions of the sprung member, and one sprung acceleration sensor is abnormal, the two front wheel side suspension devices and the two Among the rear wheel suspension devices, a three-wheel model is used in which the suspension device on the side where the abnormal sensor is installed is replaced with one virtual suspension device. Thereby, even when one of the sensors fails, the control of the damping force is continued.

  The three-wheel model may be designed so that the virtual suspension device is attached to the sprung member at an intermediate position between the two suspension devices to be replaced are attached to the sprung member. According to this, the suspension device at each wheel position is based on a three-wheel model in which the virtual suspension device is attached to an intermediate position between the positions where the two suspension devices to be replaced with the virtual suspension device are attached to the sprung member. By controlling the damping force, the imbalance between the left and right damping forces is corrected. Further, when the required damping force calculated for the virtual suspension device is distributed to the two replaced suspension devices, the distribution ratio is preferably 1: 1.

1 is an overall schematic diagram of a suspension control device for a vehicle. It is the figure which showed the suspension apparatus typically. It is a figure showing roughly the electric control device concerning a 1st embodiment. It is the figure which divided and represented suspension ECU which concerns on 1st Embodiment for every function. It is a flowchart which shows the flow of the rear-wheel approximation control program which concerns on 1st Embodiment. It is a flowchart which shows the flow of the front wheel approximate control program which concerns on 1st Embodiment. It is a flowchart which shows the flow of the turning state determination program which concerns on 1st Embodiment. It is a flowchart which shows the flow of the integrated control program which concerns on 1st Embodiment. It is a figure showing a rear-wheel approximate three-wheel model. It is the figure which represented on the sprung member the action position and acceleration of force related to the equation of motion derived from the rear wheel approximate three-wheel model. It is a block diagram of a control system (generalized plant) designed based on an equation of motion derived from a rear wheel approximate three-wheel model. It is a figure showing a front wheel approximate three-wheel model. It is the figure which represented on the sprung member the action position and acceleration of force related to the equation of motion derived from the front wheel approximate three-wheel model. It is a graph showing the relationship between the actual yaw rate and the target yaw rate and the turning state of the vehicle. It is a figure which shows schematically the electric control apparatus which concerns on 2nd Embodiment. It is the figure which divided and represented suspension ECU which concerns on 2nd Embodiment for every function. It is a flowchart which shows the flow of the integrated control program which concerns on 2nd Embodiment. It is a figure which shows schematically the electric control apparatus which concerns on 3rd Embodiment. It is a flowchart which shows the flow of the rear-wheel approximation control program which concerns on 3rd Embodiment. It is a figure which shows schematically the electric control apparatus which concerns on 4th Embodiment. It is a flowchart which shows the flow of the front wheel approximate control program which concerns on 4th Embodiment.

(First embodiment)
Hereinafter, a first embodiment of the present invention will be described with reference to the drawings. FIG. 1 is a schematic diagram showing the entire vehicle suspension control apparatus 1 according to the present embodiment. The suspension control device 1 includes a right front wheel side suspension device 10FR, a left front wheel side suspension device 10FL, a right rear wheel side suspension device 10RR, a left rear wheel side suspension device 10RL, and each suspension device 10FR, 10FL, 10RR, The electric control apparatus 20 which controls the action | operation of 10RL is provided. The electric control device 20 corresponds to the damping force control device of the present invention.

  The right front wheel side suspension device 10FR is interposed between the sprung member HA including the vehicle body and the right front wheel WFR, and on one end side (upper end side) of the sprung member HA on the right front side (right front wheel position). It is connected and connected to the right front wheel WFR via the unsprung member LA on the other end side (lower end side). The left front wheel side suspension device 10FL is interposed between the sprung member HA and the left front wheel WFL, and is connected to the left front side (left front wheel position) of the sprung member HA on one end side and on the other end side. It is connected to the left front wheel WFL via an unsprung member LA. The right rear wheel side suspension device 10RR is interposed between the sprung member HA and the right rear wheel WRR, and is connected to the right rear side (right rear wheel position) of the sprung member HA on one end side and the other end. On the side, it is connected to the right rear wheel WRR via an unsprung member LA. The left rear wheel side suspension device 10RL is interposed between the sprung member HA and the left rear wheel WRL, and is connected to the left rear side (left rear wheel position) of the sprung member HA on one end side and the other end. On the side, it is connected to the left rear wheel WRL via an unsprung member LA. In this specification, when the suspension devices 10FR, 10FL, 10RR, and 10RL and the wheels WFR, WFL, WRR, and WRL are collectively referred to, they may be simply referred to as the suspension device 10 and the wheels W.

  As shown in FIG. 1, the suspension apparatus 10 includes a spring 11 and a damper (shock absorber) 12. The spring 11 and the damper 12 are connected to the sprung member HA at one end (upper end) and connected to the unsprung member LA at the other end. The spring 11 is configured by a metal coil spring in this embodiment, but an air spring such as an air suspension device may be used. The unsprung member LA includes a knuckle connected to the wheel W including the tire, a lower arm having one end connected to the knuckle, and the like, and supports the suspension device 10. The sprung member HA is a member supported by the suspension device 10 and includes the vehicle body.

  The damper 12 includes a cylinder 121, a piston 122, and a piston rod 123. The cylinder 121 is a cylindrical member in which a viscous fluid (for example, oil) is sealed, and is connected to an unsprung member LA (specifically, a lower arm) at a lower end thereof. The piston 122 is disposed in the cylinder 121. The piston 122 divides the inside of the cylinder 121 into an upper chamber R1 and a lower chamber R2. The piston 122 can move in the cylinder 121 in the axial direction. The piston rod 123 is connected to the piston 122 at its lower end and is connected to the sprung member HA at its upper end. The piston 122 has a communication passage that communicates the upper chamber R1 and the lower chamber R2.

  In the damper 12 configured in this way, when the sprung member HA is relatively displaced with respect to the road surface due to the wheels W getting over the road surface unevenness, the piston 122 connected to the sprung member HA side is not moved to the unsprung member. The cylinder 121 connected to the LA side (road surface side) is relatively displaced. Viscous resistance is generated by the viscous fluid flowing through the communication path formed in the piston 122 with the relative displacement. The viscous resistance attenuates the relative displacement of the sprung member HA with respect to the road surface, that is, the vibration of the sprung member HA.

  As shown in the drawing, the spring 11 is connected to a retainer attached to the outer periphery of the cylinder 121 at its lower end, and is connected to the sprung member HA at its upper end. The spring 11 generates an elastic force accompanying the relative displacement of the sprung member HA with respect to the road surface.

  FIG. 2 is a diagram schematically showing the suspension device 10 according to the present embodiment. As shown in FIG. 2, a variable throttle mechanism 13 is attached to the suspension device 10. The variable throttle mechanism 13 includes a valve 131 and an actuator 132. The valve 131 is provided in a communication passage 124 formed in the piston 122, and changes the size of at least a part of the flow path cross-sectional area of the communication passage 124, that is, the valve opening degree OP by a known throttle mechanism. The actuator 132 can be configured by, for example, a stepping motor. FIG. 1 shows actuators 132FR, 132FL, 132RR, 132RL attached to the suspension devices 10FR, 10FL, 10RR, 10RL. These actuators are disposed above the dampers 12 of the suspension devices 10FR, 10FL, 10RR, and 10RL, and are fixed to the sprung member HA. The actuator 132 is connected to the valve 131 by a control rod or the like disposed inside the piston rod 123, for example. Therefore, when the actuator 132 is operated, the valve 131 is operated accordingly, and the valve opening degree OP is changed. The flow passage cross-sectional area of the communication passage 124 is changed by changing the valve opening OP. As a result, the resistance force when the viscous fluid flows in the communication path 124 is also changed. By changing the resistance force, the damping coefficient (damping force characteristic) representing the magnitude of the damping force generated by the damper 12 is changed.

  Further, as shown in FIG. 1, the right front wheel side suspension device 10FR and the left front wheel side suspension device 10FL are connected by a front wheel side stabilizer 14. Further, the right rear wheel side suspension device 10RR and the left rear wheel side suspension device 10RL are connected by a rear wheel side stabilizer 15. The front wheel side stabilizer 14 and the rear wheel side stabilizer 15 are respectively a stabilizer bar 14a, 15a extending along the left-right direction of the vehicle, and a pair of stabilizer arms continuously extending from both ends of the stabilizer bar 14a, 15a. 14b and 15b. The stabilizer bars 14a and 15a are supported by a sprung member HA (specifically, a vehicle body) so as to be rotatable around its axis. The stabilizer arms 14b and 15b are bent forward of the vehicle from the stabilizer bars 14a and 15a, and are connected to an unsprung member LA (specifically, a lower arm) at the front ends thereof. The front wheel side stabilizer 14 and the rear wheel side stabilizer 15 provided in this way have a function of generating an anti-roll moment that cancels a roll moment generated when the vehicle turns, for example, and reducing the roll moment acting on the vehicle by the anti-roll moment. Have.

  FIG. 3 is a diagram schematically showing a connection configuration of the electric control device 20. As shown in FIG. 3, the electric control device 20 includes a suspension electronic control unit (hereinafter referred to as a suspension ECU) 21, sprung acceleration sensors 221FR, 221FL, 221RR, 221RL, and road surface vertical acceleration sensors 222FR, 222FL, 222RR, 222RL, stroke sensors 223FR, 223FL, 223RR, 223RL, roll angular acceleration sensor 224, pitch angular acceleration sensor 225, vehicle speed sensor 226, rudder angle sensor 227, yaw rate sensor 228, and drive circuits 23FR, 23FL, 23RR, 23RL are provided.

The sprung acceleration sensors 221FR, 221FL, 221RR, 221RL are attached to the respective wheel positions (right front wheel position, left front wheel position, right rear wheel position, left rear wheel position) of the sprung member HA, and springs at the positions. Right front wheel side sprung acceleration x _{b_fr} ", left rear wheel side sprung acceleration x _{b_fl} ", right rear wheel side sprung acceleration x _{b_rr} ", left rear wheel side sprung acceleration x _{b_rl} "is detected respectively. The road surface vertical acceleration sensors 222FR, 222FL, 222RR, and 222RL are attached to unsprung members LA connected to the wheels W, and by measuring the acceleration along the vertical direction of the unsprung members LA, Right front wheel side road surface acceleration x _{r_fr} ", left front wheel side road surface acceleration x _{r_fl} ", right rear wheel side road surface acceleration x _{r_rr} ", left acceleration which is the acceleration along the vertical direction of the ground road surface of the wheel W to which the member LA is connected The rear wheel side road surface acceleration _{xr_rl} "is detected. Stroke sensors 223FR, 223FL, 223RR, and 223RL are attached to each suspension device 10 and measure the displacement amount of the suspension device 10, that is, the relative displacement amount of the piston 122 with respect to the cylinder 121 of the damper 12, thereby measuring the displacement relative to the road surface. The amount of displacement of the sprung member HA, that is, the amount of relative displacement between the spring on the right front wheel and the road surface (x _{r_fr} -x _{b_fr} ), the amount of relative displacement between the left front wheel on the spring and the road surface (x _{r_fl} -x _{b_fl} ) The relative displacement amount between the side spring on the road surface (x _{r_rr} -x _{b_rr} ) and the relative displacement amount between the left rear wheel side spring on the road surface (x _{r_rl} -x _{b_rl} ) are detected. The right front wheel side sprung-road relative displacement amount (x _{r_fr} -x _{b_fr} ) is the right front wheel side sprung displacement amount x which is a displacement amount from the reference position along the vertical direction of the sprung member HA at the right front wheel position. and _{B_fr,} represented by the difference between the right front wheel side road surface displacement x _{R_fr} a displacement from a reference position along the vertical direction of the ground road surface of the right front wheel WFR. Similarly, the amount of relative displacement between the left front wheel side sprung and road surface (x _{r_fl} -x _{b_fl} ) is the difference between the left front wheel side sprung displacement amount x _{b_fl} and the left front wheel side road surface displacement amount x _{r_fl.} − Relative displacement between the road surfaces (x _{r_rr} -x _{b_rr} ) is the difference between the right rear wheel-side sprung displacement x _{b_rr} and the right rear wheel-side road displacement x _{r_rr} , the left rear wheel-side spring-to-road relative displacement (x _{r_rl} -x _{b_rl} ) is expressed by the difference between the left rear wheel-side sprung displacement amount x _{b_rl} and the left rear wheel-side road surface displacement amount x _{r_rl} , respectively.

The roll angular acceleration sensor 224 is attached to the sprung member HA, and the angular acceleration (roll angular acceleration) θ _r ″ of the roll angle θ _r that represents the angular displacement of the sprung member HA in the roll direction (around the front-rear axis). The pitch angular acceleration sensor 225 is also attached to the sprung member HA, and detects the angular acceleration (pitch angular acceleration) θ _p ”of the pitch angle θ _p representing the angular displacement of the sprung member HA around the left-right axis. To do. The vehicle speed sensor 226 is attached in the vicinity of the wheel W, and detects the vehicle speed V by counting vehicle speed pulses. The steering angle sensor 227 is attached to a steering shaft connected to the steering handle, and detects the steering angle φ of the steered wheel (generally the front wheel) by measuring the rotation angle of the steering shaft. The yaw rate sensor 228 is attached to the sprung member HA, and detects an actual yaw rate YR representing the angular velocity around the vertical axis of the sprung member HA.

  The suspension ECU 21 is a microcomputer whose main components are a CPU, a ROM, a RAM, and the like. The various sensors described above are connected to the input side of the suspension ECU 21, and detection signals from these sensors are input. The suspension ECU 21 outputs a drive signal for controlling the drive of the actuator 132 by executing various programs including a program described later based on detection signals from the various sensors. Thereby, the damping force generated by each damper 12 of the suspension device 10 is controlled.

  The drive circuits 23FR, 23FL, 23RR, 23RL are connected to the output side of the suspension ECU 21. The drive circuits 23FR, 23FL, 23RR, and 23RL are connected to the actuators 132FR, 132FL, 132RR, and 132RL corresponding to the suspension devices 10FR, 10FL, 10RR, and 10RL, respectively, and are based on the drive signals output from the suspension ECU 21. The drive current is output to the actuators 132FR, 132FL, 132RR, 132RL.

FIG. 4 shows the suspension ECU 21 divided into functions. As shown in FIG. 4, the suspension ECU 21 includes a rear wheel approximation control unit 211, a front wheel approximation control unit 212, a turning state determination unit 213, and an integrated control unit 214. The rear wheel approximate control unit 211 is a first right front wheel side request that is a damping force of a control target that should be generated by the dampers of the suspension devices 10FR, 10FL, 10RR, and 10RL, based on a rear wheel approximate three-wheel model that will be described later. The damping force F1 _{req_fr} , the first left front wheel side requested damping force F1 _{req_fl} , the first right rear wheel side requested damping force F1 _{req_rr} , and the first left rear wheel side requested damping force F1 _{req_rl} are calculated, and these calculated requested damping forces are calculated. Is output.

The front wheel approximate control unit 212 is a second right front wheel side required damping force that is a damping force of a control target that should be generated by the dampers of the suspension devices 10FR, 10FL, 10RR, and 10RL based on a front wheel approximate three-wheel model to be described later. F2 _{req_fr} , second left front wheel side required damping force F2 _{req_fl} , second right rear wheel side required damping force F2 _{req_rr} , second left rear wheel side required damping force F2 _{req_rl} are calculated, and these calculated required damping forces are output. To do.

  The turning state determination unit 213 determines whether the current turning state of the vehicle, specifically, the current turning state of the vehicle is an understeer state or an oversteer state, by executing a turning state determination program described later. Then, the turning state flag RS representing the determination result is output.

The integrated control unit 214 receives the first right front wheel side required damping force F1 _{req_fr} , the first left front wheel side required damping force F1 _{req_fl} , the first right rear wheel side required damping force F1 _{req_rr} , and the first left from the rear wheel approximation control unit 211. The rear wheel side required damping force F1 _{req_rl} is _obtained from the front wheel approximate control unit 212 by the second right front wheel side required damping force F2 _{req_fr} , the second left front wheel side required damping force F2 _{req_fl} , the second right rear wheel side required damping force F2 _{req_rr} , The second left rear wheel side required damping force F2 _{req_rl} is input from the turning state determination unit 213 to the turning state flag RS. Further, based on the input value, the right front wheel side required damping force F _{req_fr} that is a damping force that should be generated by the damper of the right front wheel side suspension device 10FR, and the damping force that should be generated by the damper of the left front wheel side suspension device 10FL. Left front wheel side required damping force F _{req_fl} , right front wheel side required damping force F _{req_rr} , which is a damping force that should be generated by the damper of the right rear wheel side suspension device _{10 RR} , and damping force that should be generated by the damper of the left rear wheel side suspension device _{10 RL} The left rear wheel side required damping force F _{req_rl} is finally determined (selected). Then, the drive signal corresponding to the determined right front wheel side required damping force F _{req_fr} is _sent to the drive circuit 23FR, the drive signal corresponding to the determined left front wheel side required damping force F _{req_fl} is _sent to the drive circuit 23FL, and the determined right rear wheel side. A drive signal corresponding to the required damping force F _{req_rl} is output to the drive circuit 23RR, and a drive signal corresponding to the determined left rear wheel side required damping force F _{req_rl} is output to the drive circuit 23RL.

  In the suspension control device 1 configured as described above, when any one of the sprung accelerations at the respective wheel positions obtained from the detection values of the sprung acceleration sensors 221FR, 221FL, 221RR, 221RL exceeds a predetermined threshold value, the suspension ECU 21 The wheel approximate control unit 211 executes a rear wheel approximate control program, the front wheel approximate control unit 212 executes a front wheel approximate control program, the turning state determination unit 213 executes a turning state determination program, and the integrated control unit 214 executes an integrated control program. .

FIG. 5 is a flowchart showing the flow of the rear wheel approximation control program. The rear wheel approximation control unit 211 starts this rear wheel approximation control program at step (hereinafter, step number is abbreviated as S) 100 in FIG. Next, at S102, the sprung acceleration sensors 221FR, 221FL, 221RR, and 221RL, the right front wheel side sprung acceleration x _{b_fr} ", the left front wheel side sprung acceleration x _{b_fl} ", the right rear wheel side sprung acceleration x _{b_rr} ", left The rear wheel side sprung acceleration x _{b_rl} "is determined from the road surface vertical acceleration sensors 222FR, 222FL, 222RR and 222RL, the right front wheel side road surface acceleration _{xr_fr} ", the left front wheel side road surface acceleration _{xr_fl} ", and the right rear wheel side road surface acceleration _{xr_rr} ". , Left rear wheel side road surface acceleration x _{r_rl} "is determined from the stroke sensors 223FR, 223FL, 223RR, 223RL, the amount of relative displacement between the right front wheel side spring on the road surface (x _{r_fr} -x _{b_fr} ), the left front wheel side spring on the road surface relative to the road surface. Displacement amount (x _{r_fl} -x _{b_fl} ), right rear wheel side sprung-road relative displacement amount (x _{r_rr} -x _{b_rr} ), left rear wheel side sprung-road relative displacement amount (x _{r_rl} -x _{b_rl} ) Roll angular acceleration from roll angular acceleration sensor 224 "The pitch angle acceleration theta _p from the pitch angular acceleration sensor 225" theta _r, and inputs respectively.

Next, in S104, right front wheel side sprung acceleration x _{b_fr} ”is integrated over time to obtain right front wheel side sprung speed x _{b_fr} ′ which is the speed along the vertical direction of sprung member HA at the right front wheel position. , calculates the left front wheel on the side spring rate _x b_fl 'by time integration of the acceleration _x b_fl "on the left front wheel-side spring. Also, right rear wheel side sprung acceleration x _{b_rr} "is time integrated to right rear wheel side sprung speed x _{b_rr} 'and left rear wheel side sprung acceleration x _{b_rl} " is time integrated to left rear wheel side sprung 'calculates the further these speeds x _{B_rr'} rate x _{B_rl} 'by dividing by 2 by adding the velocity on the rear center wheel spring x _{B_R'} and x _{B_rl} calculated.

Further, the right front wheel side sprung speed x _{b_fr} ′ is further integrated over time, so that the right front wheel side sprung displacement x _{b_fr} which is a displacement amount from the reference position along the vertical direction of the sprung member HA at the right front wheel position. The left front wheel side sprung speed x _{b_fl} ′ is further integrated over time to calculate the left front wheel side sprung displacement x _{b_fl} . Further, the rear center wheel side sprung speed x _{b_R} ′ is further integrated over time to calculate the rear center wheel side sprung displacement x _{b_R} .

Subsequently, at S106, the right front wheel side road surface acceleration _{xr_fr} "is time-integrated to obtain the right front wheel side road surface speed _{xr_fr} 'which is the displacement speed along the vertical direction of the ground contact road surface of the right front wheel WFR. The left front wheel side road surface speed x _{r_fl} ′ is calculated by integrating the side road surface acceleration x _{r_fl} ”over time. In addition, the right rear wheel side road surface acceleration x _{r_rr} "is time-integrated to obtain the right rear wheel side road surface speed x _{r_rr} ', and the left rear wheel side road surface acceleration x _{r_rl} " is time integrated to obtain the left rear wheel side road surface speed x _{r_rl} '. was calculated, further, by dividing by 2 by adding these velocity _x r_rr 'and _x r_rl', calculates a rear center wheel side road surface velocity x _{r_r} '.

Further, the right front wheel side road surface speed x _{r_fr} ′ is further integrated over time, so that the right front wheel side road surface displacement amount x _{r_fr} , which is a displacement amount from the reference position along the vertical direction of the ground road surface of the right front wheel WFR, is _converted into the left front wheel. The left front wheel side road surface displacement x _{b_fl} is calculated by further integrating the side road surface speed x _{r_fl} ′ over time. Further, the rear center wheel side road surface displacement _{xr_R} ′ is further time integrated to calculate the rear center wheel side road surface displacement amount _{xr_R} .

Next, at S108, the right front wheel side sprung speed x _{b_fr} ′ and the right front wheel side road speed x _{r_fr} ′ are obtained by time differentiation of the relative displacement (x _{r_fr} −x _{b_fr} ) between the right front wheel sprung and road surface. The right front wheel side sprung-road relative speed (x _{r_fr} '-x _{b_fr} ') and the left front wheel side sprung-road relative displacement (x _{r_fl} -x _{b_fl} ) _expressed by the difference The left front wheel side sprung-road relative speed (x _{r_fl} '-x _{b_fl} ') is calculated. Also, the right rear wheel side sprung-road relative displacement (x _{r_rr} -x _{b_rr} ) is time-differentiated to obtain the right rear wheel side sprung-road relative speed (x _{r_rr} '-x _{b_rr} '). Relative displacement amount between wheel side sprung and road surface (x _{r_rl} -x _{b_rl} ) is time differentiated to _calculate left rear wheel side sprung-road relative speed (x _{r_rl} '-x _{b_rl} '). _{Add the} speed (x _{r_rr} '-x _{b_rr} ') and (x _{r_rl} '-x _{b_rl} ') and divide by 2 to obtain the rear center wheel side sprung-road relative speed (x _{r_R} '-x _{b_R} ') calculate. This rear center wheel side sprung-road relative speed (x _{r_R} '-x _{b_R} ') is the difference between the rear center wheel side sprung speed x _{b_R} 'and the rear center wheel side road surface speed x _{r_R} '.

Next, in S110, the rear wheel approximation control unit 211 applies a nonlinear H _∞ control theory to the right front wheel side variable damping coefficient C _{v_fr} , the left front wheel side variable damping coefficient C _{v_fl} , and the rear center wheel side variable damping coefficient C. _{v_R} is calculated. In performing this calculation, the rear wheel approximate three-wheel model shown in FIG. 9 is used as a dynamic motion model. In this rear wheel approximate three-wheel model, the right rear wheel side suspension device 10RR and the left rear wheel side suspension device 10RL are replaced with one rear center wheel side suspension device 10R, and the sprung member HA is located on the front wheel side. This is a vehicle model representing the motion of a vehicle supported by two suspension devices 10FR and 10FL and one suspension device 10R located at the center of the rear wheel. The rear center wheel suspension device 10R is virtually attached to the sprung member HA at the center position of the rear wheel shaft connecting the right rear wheel position and the left rear wheel position of the sprung member HA. The rear center wheel side sprung speed x _{b_R} ′ and the rear center wheel side sprung displacement x _{b_R} calculated in S104 are positions where the rear center wheel side suspension device 10R is attached to the sprung member HA ( This is the speed and displacement along the vertical direction of the sprung member HA at the rear center wheel position). Further, the rear center wheel side road surface speed x _{r_R} ′ and the rear center wheel side road surface displacement amount x _{r_R} calculated in S106 are the ground road surface of the wheel (rear center wheel) to be connected to the rear center wheel side suspension device 10R. The speed and displacement along the vertical direction. The rear central wheel side suspension device 10R corresponds to the virtual suspension device and the rear wheel side virtual suspension device of the present invention.

Further, in the rear wheel approximate three-wheel model of FIG. 9, the spring constant of the spring of the right front wheel side suspension device 10FR is K _{s_fr and} the spring constant of the spring of the left front wheel side suspension device 10FL is K _{s_fl.} The spring constant of the spring of the device 10R is represented by K _{s_R} . The spring constant K _{s_R} can be expressed, for example, by the average of the spring constant K _{s_rr} of the spring of the right rear wheel suspension apparatus 10RR and the spring constant K _{s_rl} of the spring of the left rear wheel suspension apparatus 10RL.

Further, the damping coefficient of the damper of the right front wheel side suspension device 10FR is represented by the sum (C _{s_fr} + C _{v_fr} ) of the right front wheel side linear damping coefficient C _{s_fr} and the right front wheel side variable damping coefficient C _{v_fr} . The right front wheel side linear damping coefficient C _{s_fr} represents a damping coefficient that does not vary, and the right front wheel side variable damping coefficient C _{v_fr} represents a varying damping coefficient. Further, the damping coefficient of the damper of the left front wheel side suspension apparatus 10FL is represented by the sum (C _{s_fl} + C _{v_fl} ) of the left front wheel side linear damping coefficient C _{s_fl} and the left front wheel side variable damping coefficient C _{v_fl} . The left front wheel side linear damping coefficient C _{s_fl} represents a damping coefficient that does not vary, and the left front wheel side variable damping coefficient C _{v_fl} represents a varying damping coefficient. Further, the damping coefficient of the damper of the rear center wheel side suspension apparatus 10R is represented by the sum (C _{s_R} + C _{v_R} ) of the rear center wheel side linear damping coefficient C _{s_R} and the rear center wheel side variable damping coefficient C _{v_R} . The rear center wheel side linear damping coefficient C _{s_R} represents a damping coefficient that does not vary, and the rear center wheel side variable damping coefficient C _{v_R} represents a varying damping coefficient. Each linear damping coefficient C _{s_fr} , C _{s_fl} , C _{s_R} is predetermined. Rear center wheel linear damping coefficient C _{S_R,} for example after the left defined in advance for the right rear wheel suspension system 10RR predefined be right rear wheel linear damping coefficient C _{S_rr} and the left rear wheel suspension system 10RL damper for the damper It can be represented by the average of the ring-side linear damping coefficient C _{s_rl} . Further, a front wheel side torsional elastic coefficient that is a coefficient of torsional force generated by the front wheel side stabilizer 14 is represented by K _{stb_F} .

In S110, the rear wheel approximate control unit 211 applies the nonlinear _H∞ control theory to the control system designed based on the equation of motion derived from the rear wheel approximate three-wheel model, and is represented in the model. The variable damping coefficients (right front wheel side variable damping coefficient C _{v_fr} , left front wheel side variable damping coefficient C _{v_fl} , rear center wheel side variable damping coefficient C _{v_R} ) of the three suspension devices 10FR, 10FL, 10R are calculated.

The equation of motion of the sprung member HA derived from the rear wheel approximate three-wheel model is expressed by the following equations (eq.1) to (eq.3).

Equation (eq.1) is the equation of motion of the sprung member HA in the vertical direction (heave equation of motion), and equation (eq.2) is the equation of motion of the sprung member HA in the rolling direction (rotational direction around the longitudinal axis). (Roll motion equation), equation (eq.3) is a motion equation (pitch motion equation) of the sprung member HA in the pitching direction (rotational direction around the left-right axis). In the above equation, M _b is the mass of the sprung member HA, x _h ″ is the acceleration along the vertical direction (heave acceleration) at the center of gravity of the sprung member HA. Heave acceleration x _h ″ is each sprung acceleration. It can be obtained from the sprung acceleration detected by the sensors 221FR, 221FL, 221RR, 221RL. F _{sus_fr} is a force acting vertically on the sprung member HA at the right front wheel position of the sprung member HA (right front wheel side vertical force), and F _{sus_fl} is applied to the sprung member HA at the left front wheel position of the sprung member HA. The force acting in the vertical direction (left front wheel side vertical force), F _{sus_R} is the force acting in the vertical direction on the sprung member HA at the rear central wheel position of the sprung member HA (rear central wheel side vertical force). FIG. 10 is a diagram showing the applied position and acceleration of the vertical force expressed by the equations (eq.1) to (eq.3).

Further, in the above equation, I _r is the moment of inertia (roll inertia) of the rolling direction of the sprung member HA, pitching direction of the inertia moment (pitch inertia) of I _p is sprung member HA, T _f is a front wheel tread is there. L _f is a longitudinal distance from the center of gravity position of the sprung member HA to the front wheel axis connecting the right front wheel position and the left front wheel position. L _r is the distance in the front-rear direction from the position of the center of gravity of the sprung member HA to the position of the rear center wheel.

The right front wheel side vertical force F _{sus_fr} is expressed by the following equation (eq.4), the left front wheel side vertical force F _{sus_fl} is expressed by the following equation (eq.5), and the rear center wheel side vertical force F _{sus_R} is _expressed by the following equation (eq.6). Is represented by

In the above equations (eq.4) to (eq.6), F _{sp_fr} is the elastic force (right front wheel side elastic force) generated by the spring of the right front wheel side suspension device 10FR, and F _{sp_fl} is the left front wheel side suspension device 10FL. The elastic force (left front wheel side elastic force) generated by the spring and F _{sp_R} is the elastic force (rear center wheel side elastic force) generated by the spring of the rear central wheel side suspension device 10R. F _{Cs_fr} is the damping force (right front wheel side linear damping force) represented by the right front wheel side linear damping coefficient C _{s_fr} among the damping forces generated by the damper of the right front wheel side suspension device 10FR, and F _{Cv_fr} is the right front wheel side variable. This is the damping force (right front wheel side variable damping force) represented by the damping coefficient _{Cv_fr} . F _{Cs_fl} is the damping force (left front wheel side linear damping force) represented by the left front wheel side linear damping coefficient C _{s_fl} of the damping force generated by the damper of the left front wheel side suspension device 10FL, and F _{Cv_fl} is the left front wheel side variable This is the damping force (left front wheel side variable damping force) represented by the damping coefficient C _{v_fl} . F _{Cs_R} is the damping force (rear center wheel side linear damping force) represented by the rear center wheel side linear damping coefficient C _{s_R} among the damping forces generated by the damper of the rear center wheel side suspension device 10R, and F _{Cv_R} is the rear side. This is the damping force (rear center wheel side variable damping force) represented by the center wheel side variable damping coefficient _{Cv_R} . F _{stb_F} is a torsional elastic force (front wheel side torsional elastic force) generated by the front wheel side stabilizer 14.

The right front wheel side elastic force F _{sp_fr} , the left front wheel side elastic force F _{sp_fl} and the rear center wheel side elastic force F _{sp_R} are represented by the following equations (eq.7) to (eq.9).

The right front wheel side linear damping force F _{Cs_fr} , the left front wheel side linear damping force F _{Cs_fl} and the rear center wheel side linear damping force F _{Cs_R} are expressed by the following equations (eq.10) to (eq.12).

The right front wheel side variable damping force F _{Cv_fr} , the left front wheel side variable damping force F _{Cv_fl} and the rear center wheel side variable damping force F _{Cv_R} are represented by the following equations (eq.13) to (eq.15).

The front wheel side torsional elastic force F _{stb_F} is expressed by the following equation (eq.16).

Further, the heave displacement amount x _h , the roll angle θ _r and the pitch angle θ _p representing the vertical displacement amount at the center of gravity position of the sprung member HA use a mode conversion matrix as shown in the following equation (eq.17). Thus, the right front wheel side sprung displacement amount x _{b_fr} , the left front wheel side sprung displacement amount x _{b_fl} , and the sprung displacement amount at the rear center wheel position (front center wheel side sprung displacement amount) x _{b_R} can be converted.

From the above relational expression, the state space representation of the rear wheel approximate three-wheel model is expressed as the following expression (eq.18).

In the above equation (eq.18), the state quantity x _p , the evaluation output z _p , and the disturbance w are expressed as, for example, the following equation (eq.19).

Further, the control input u is a right front wheel side variable damping coefficient C _{v_fr} , a left front wheel side variable damping coefficient C _{v_fl} and a rear center wheel side variable damping coefficient C _{v_R} as shown in the following equation (eq.20).

A _p , B _P1 , B _p2 (x _p ), C _p1 , and D _p12 (x _p ) are coefficient matrices. These coefficient matrices are obtained when the state quantity x _p , the evaluation output z _p , and the disturbance w are set as in (eq.19) above, and the control input u is set as in the above equation (eq.20). Are defined to satisfy the equations of motion of equations (eq.1) to (eq.3).

Table 2 shows the number of manipulated variables (control input u) in the state space expression shown in equation (eq.18), the number of state quantities, and the degree of freedom of the rear wheel approximate three-wheel model that is the basis of the state space expression. Indicates.

As shown in Table 2, the degree of freedom of the sprung member represented by the rear wheel approximate three-wheel model is three degrees of freedom (vertical motion, roll motion, pitch motion). Further, in this model, the movement of the unsprung member is not considered. That is, the unsprung member is displaced up and down like the road surface. For this reason, the degree of freedom of the suspension device is one degree of freedom (up-and-down motion of the sprung member). Therefore, this model has 3 degrees of freedom. Further, the disturbance of the system is the road surface displacement amounts x _{r_fr} , x _{r_fl} , x _{r_R} of the three wheels (the right front wheel, the left front wheel, and the rear center wheel) as shown in the equation (eq.19). When this disturbance is mode-converted into the displacement of the sprung member HA, the disturbance is expressed by the vertical displacement, roll displacement, and pitch displacement of the sprung member HA. Therefore, the degree of freedom of the sprung member HA due to disturbance is 3 degrees of freedom. Since the degree of freedom of the model and the degree of freedom of the sprung member HA due to disturbance are equal, this control system is controllable.

FIG. 11 is a block diagram of a generalized plant (control system) designed based on the state space expression shown in the above equation (eq.18). As shown in the figure, frequency weights W _s (s) and W _u (s) act on the evaluation output z _p and the control input u, respectively. It should be noted that the system may be designed so that a nonlinear weight further acts on this output.

上式において、x_w'は状態量x_wの微分を表し、x_u'は状態量x_uの微分を表す。 State-space representation of the frequency weight W _s (s) the state quantity x _w of frequency weight W _s (s), the output z _w and the constant matrix _{A w, B w, C w} , the D _w, the following equation (eq .21). The state space representation of the frequency weight W _u (s) is the state amount x _u of the frequency weight W _u (s), the output z _u and the constant matrix _{A u, B u, C u} , the D _u, the following equation It is expressed as (eq.22).

In the above equation, x _w ′ represents the derivative of the state quantity x _w , and x _u ′ represents the derivative of the state quantity x _u .

The state space representation of the generalized plant can be expressed as the following equation (eq.23) by using the equations (eq.18), (eq.21), and (eq.22).

Equation (eq.23) is a bilinear system. Therefore, the control law u = k (x) of this system is obtained by solving the Riccati inequality approximately shown in the following equation (eq.24) instead of solving the Hamilton Jacobi partial differential inequality.

When there is a positive definite symmetric matrix P satisfying the Riccati inequality represented by the above equation (eq.24), the closed-loop system of the generalized plant shown in FIG. 11 is internally stable and L2 representing robustness against the disturbance w The gain is a positive constant γ or less. At this time, the control input u (= k (x)) is expressed, for example, by the following equation (eq.25).

Therefore, in S110, the rear wheel approximation control unit 211 can obtain the variable damping coefficients C _{v_fr} , C _{v_fl} , C _{v_R} by obtaining the control input u based on the equation (eq.25).

After calculating the variable damping coefficient, the rear wheel approximation control unit 211 proceeds to S112 in FIG. 5, and the first right front wheel side required damping force F1 _{req_fr} , the first left front wheel side required damping force F1 _{req_fl} , the first rear center wheel side. Calculate the required damping force F1 _{req_R} . These required damping forces are damping forces (required damping forces) of the control target that should be generated by the dampers of the respective suspension devices represented in the rear wheel approximate three-wheel model. The first right front wheel side required damping force F1 _{req_fr} is a required damping force for the right front wheel side suspension device 10FR represented by the rear wheel approximate three-wheel model, the right front wheel side linear damping coefficient C _{s_fr} and the right front wheel side variable damping It is calculated by multiplying the sum (C _{s_fr} + C _{v_fr} ) with the coefficient C _{v_fr} by the right front wheel side sprung-road relative speed (x _{r_fr} '-x _{b_fr} '). First left front wheel requested damping force F1 _{Req_fl} is demanded damping force for the left front wheel suspension system 10FL represented to the rear wheel approximate three-wheeled model, the left front wheel linear damping coefficient C _{S_fl} and left front side variable attenuation It is calculated by multiplying the sum (C _{s_fl} + C _{v_fl} ) with the coefficient C _{v_fl} by the left front wheel side sprung-road relative speed (x _{r_fl} '-x _{b_fl} '). The first rear center wheel demanded damping force F1 _{req_r} is demanded damping force for the rear center wheel suspension system 10R represented to the rear wheel approximate three-wheeled model, the rear center wheel linear damping coefficient C _{S_R} and the rear center It is calculated by multiplying the sum (C _{s_R} + C _{v_R} ) with the wheel-side variable damping coefficient C _{v_R} by the rear center wheel-side sprung-road relative speed (x _{r_R} '-x _{b_R} '). The processing for calculating the required damping force for the three suspension devices 10FR, 10FL, 10R represented in the rear wheel approximate three-wheel model in S112 corresponds to the model required damping force calculation means of the present invention.

Next, in S114, the rear wheel approximation control unit 211 _changes the first rear center wheel side required damping force F1 _{req_R} to the first right rear wheel side required damping force F1 _{req_rr} and the first left rear wheel side required damping force F1 _{req_rl.} To distribute. Specifically, the first right rear wheel side required damping force F1 _{req_rr} and the first left rear wheel side required damping force F1 _{req_rl} are calculated by dividing the first rear center wheel side required damping force F1 _{req_R} by 2. The first right rear wheel side required damping force F1 _{req_rr} is a required damping force for the right rear wheel side suspension device 10RR calculated based on the rear wheel approximate three-wheel model, and the first left rear wheel side required damping force F1. _{req_rl} is a required damping force for the left rear wheel side suspension device 10RL. When the mounting position of the rear center wheel suspension device 10R is located between the mounting position of the right rear wheel suspension device 10RR and the mounting position of the left rear wheel suspension device 10RL as in the present embodiment, the first right as the size and magnitude of the first left rear wheel demanded damping force F1 _{Req_rl} the rear wheel requested damping force F1 _{Req_rr} equal, i.e. the distribution ratio of 1: to be 1, the first rear center wheel side request The damping force F1 _{req_R} is _distributed to the first right rear wheel side required damping force F1 _{req_rr} and the first left rear wheel side required damping force F1 _{req_rl} . The process performed in S114 corresponds to the required damping force distribution means of the present invention.

After calculating the first right rear wheel side required damping force F1 _{req_rr} and the first left rear wheel side required damping force F1 _{req_rl in S114} , the rear wheel approximation control unit 211 proceeds to S116, and the first right front wheel side required damping force is calculated. F1 _{req_rl} , first left front wheel side required damping force F1 _{req_rl} , first right rear wheel side required damping force F1 _{req_rl} , and first left rear wheel side required damping force F1 _{req_rl} are output. Thereafter, the process proceeds to S118 and the program is terminated. As can be seen from the above description, the rear wheel approximate control unit 211 of the present embodiment is a control designed based on the vertical motion, roll motion and pitch motion of the sprung member HA derived from the rear wheel approximate three-wheel model. By applying nonlinear _H∞ control theory to the system, the required damping force for the three suspension devices represented in the rear wheel approximate three-wheel model is calculated. Further, of the calculated requested damping force, the requested right damping force for the rear center wheel side suspension device 10R (first rear center wheel side requested damping force F1 _{req_R} ) is replaced with the rear center wheel side suspension device 10R. It distributes to the required damping force (first right rear wheel side required damping force F1 _{req_rr} and first left rear wheel side required damping force F1 _{req_rl} ) for the wheel side suspension device 10RR and the left rear wheel side suspension device 10RL. As a result, the required damping force for the four suspension devices attached to each wheel position of the sprung member HA is calculated. The rear wheel approximation control unit 211 corresponds to the first required damping force calculation means of the present invention.

FIG. 6 is a flowchart showing the flow of the front wheel approximate control program. The front wheel approximation control unit 212 starts this front wheel approximation control program in S200 of FIG. Next, in S202, the sprung acceleration sensors 221FR, 221FL, 221RR, 221RL, the right front wheel side sprung acceleration _{xb_fr} ", the left front wheel side sprung acceleration _{xb_fl} ", the right rear wheel side sprung acceleration _{xb_rr} ", left The rear wheel side sprung acceleration x _{b_rl} "is determined from the road surface vertical acceleration sensors 222FR, 222FL, 222RR and 222RL, the right front wheel side road surface acceleration _{xr_fr} ", the left front wheel side road surface acceleration _{xr_fl} ", and the right rear wheel side road surface acceleration _{xr_rr} ". , Left rear wheel side road surface acceleration x _{r_rl} "is determined from the stroke sensors 223FR, 223FL, 223RR, 223RL, the amount of relative displacement between the right front wheel side spring on the road surface (x _{r_fr} -x _{b_fr} ), the left front wheel side spring on the road surface relative to the road surface. Displacement amount (x _{r_fl} -x _{b_fl} ), right rear wheel side sprung-road relative displacement amount (x _{r_rr} -x _{b_rr} ), left rear wheel side sprung-road relative displacement amount (x _{r_rl} -x _{b_rl} ) Roll angular acceleration from roll angular acceleration sensor 224 "The pitch angle acceleration theta _p from the pitch angular acceleration sensor 225" theta _r, and inputs respectively.

Then, the left front wheel by at S204, "the speed on the front-right wheel-side spring _x b_fr 'by time integrating the acceleration x _{B_fl} on the left front wheel-side spring" on the right front wheel-side spring acceleration x _{B_fr} to time integration of the The side sprung speed x _{b_fl} ′ is calculated, and these speeds x _{b_fr} ′ and x _{b_fl} ′ are added and divided by two to calculate the front wheel center wheel side sprung speed x _{b_F} ′. Also, right rear wheel side sprung acceleration x _{b_rr} "is time integrated to right rear wheel side sprung speed x _{b_rr} 'and left rear wheel side sprung acceleration x _{b_rl} " is time integrated to left rear wheel side sprung Calculate velocity x _{b_rl} '.

Further, the right rear wheel sprung speed x _{b_rr} ′ is further integrated over time, so that the right rear wheel sprung displacement amount, which is the displacement amount from the reference position along the vertical direction of the sprung member HA at the right rear wheel position, is obtained. The left rear wheel side sprung displacement amount x _{b_rl} is calculated by further integrating x _{b_rr} with the left rear wheel side sprung speed x _{b_rl} 'over time. Further, the front center wheel side sprung speed x _{b_F} ′ is further integrated over time to calculate the front center wheel side sprung displacement x _{b_F} .

Subsequently, in S206, the right front wheel side road surface speed _{xr_fr} "is time-integrated to integrate the right front wheel side road surface speed _{xr_fr} ', and the left front wheel side road surface acceleration _{xr_fl} " is integrated to time. x _{r_fl} ′ is calculated, and these speeds x _{r_fr} ′ and x _{r_fl} ″ are added and divided by two to calculate the front center wheel side road speed x _{r_F} ′. Also, the right rear wheel side road acceleration x _{r_rr} Time integration is performed to calculate the right rear wheel side road surface speed _{xr_rr} 'and the left rear wheel side road surface acceleration _{xr_rl} "to time integration to calculate the left rear wheel side road surface speed _{xr_rl} '.

Further, the right rear wheel side road surface speed _{xr_rr} ′ is further integrated over time to obtain the right rear wheel side road surface displacement amount _{xr_rr} which is a displacement amount from the reference position along the vertical direction of the ground road surface of the right rear wheel WRR. , calculates the left rear wheel side road surface displacement x _{R_rl} by further time integration of the left rear wheel side road surface velocity _x r_rl '. Further, the front center wheel side road surface displacement x _{r_F} 'is further integrated over time to calculate the front center wheel side road surface displacement amount x _{r_F} .

Next, in S208, the right front wheel side sprung-road relative speed (x _{r_fr} '-x _{b_fr} ') is obtained by time-differentiating the right front wheel side sprung-road relative displacement amount (x _{r_fr} -x _{b_fr} ). The left front wheel side sprung-road relative speed (x _{r_fl} '-x _{b_fl} ') is calculated by differentiating the left front wheel side sprung-road relative displacement (x _{r_fl} -x _{b_fl} ) with _respect to time. By _adding these speeds (x _{r_fr} '-x _{b_fr} ') and (x _{r_fl} '-x _{b_fl} ') and dividing by 2, the relative speed between the front center wheel side sprung and road surface (x _{r_F} '-x _{b_F} ' ). This front center wheel side sprung-road relative speed (x _{r_F} '-x _{b_F} ') is the difference between the front center wheel side sprung speed x _{b_F} 'and the front center wheel side road surface speed x _{r_F} '. Also, the right rear wheel side sprung-road relative displacement (x _{r_rr} -x _{b_rr} ) is time-differentiated to obtain the right rear wheel side sprung-road relative speed (x _{r_rr} '-x _{b_rr} '). The relative displacement between the wheel-side sprung and road surface (x _{r_rl} -x _{b_rl} ) is time-differentiated to _calculate the left rear wheel-side sprung-road relative speed (x _{r_rl} '-x _{b_rl} ').

Next, in S210, the front wheel approximate control unit 212 applies the non-linear _H∞ control theory to the front center wheel side variable damping coefficient C _{v_F} , the right rear wheel side variable damping coefficient C _{v_rr} , and the left rear wheel side variable damping coefficient. C _{v_rl} is calculated. In performing this calculation, the front wheel approximate three-wheel model shown in FIG. 12 is used as a dynamic motion model. In the front wheel approximate three-wheel model, the right front wheel side suspension device 10FR and the left front wheel side suspension device 10FL are replaced with one front center wheel side suspension device 10F, and the sprung member HA is located at the center of the front wheel. This is a vehicle model representing the motion of the vehicle supported by the suspension device 10F and the two suspension devices 10RR and 10RL located on the rear wheel side. The front center wheel suspension device 10F is virtually attached to the sprung member HA at the center position of the front wheel shaft connecting the right front wheel position and the left front wheel position of the sprung member HA. The front center wheel side sprung speed x _{b_F} ′ and the front center wheel side sprung displacement amount x _{b_F} calculated in S204 are positions where the front center wheel side suspension device 10F is attached to the sprung member HA ( This is the speed and displacement along the vertical direction of the sprung member HA at the front center wheel position). Further, the front center wheel side road surface speed x _{r_F} ′ and the front center wheel side road surface displacement amount x _{r_F} calculated in S206 are the _values of the ground road surface of the wheel (front center wheel) to be connected to the front center wheel side suspension device 10F. The speed and displacement along the vertical direction. The front center wheel suspension device 10F corresponds to the virtual suspension device and the front wheel virtual suspension device of the present invention.

Further, in the front wheel approximate three-wheel model of FIG. 12, the spring constant of the spring of the right rear wheel side suspension device 10RR is K _{s_rr and} the spring constant of the spring of the left rear wheel side suspension device 10RL is K _{s_rl.} The spring constant of the suspension device 10F is represented by K _{s_F} . The spring constant K _{s_F} can be _expressed by, for example, the average of the spring constant K _{s_fr} of the spring of the right front wheel side suspension device 10FR and the spring constant K _{s_fl} of the spring of the left front wheel side suspension device 10FL.

Further, the damping coefficient of the damper of the right rear wheel side suspension apparatus 10RR is represented by the sum (C _{s_rr} + C _{v_rr} ) of the right rear wheel side linear damping coefficient C _{s_rr} and the right rear wheel side variable damping coefficient C _{v_rr} . The right rear wheel side linear damping coefficient C _{s_rr} represents a damping coefficient that does not vary, and the right rear wheel side variable damping coefficient C _{v_rr} represents a varying damping coefficient. Further, the damping coefficient of the damper of the left rear wheel side suspension apparatus 10RL is represented by the sum (C _{s_rl} + C _{v_rl} ) of the left rear wheel side linear damping coefficient C _{s_rl} and the left rear wheel side variable damping coefficient C _{v_rl} . The left rear wheel side linear damping coefficient C _{s_rl} represents a damping coefficient that does not vary, and the left rear wheel side variable damping coefficient C _{v_rl} represents a varying damping coefficient. Further, the damping coefficient of the damper of the front center wheel side suspension apparatus 10F is represented by the sum (C _{s_F} + C _{v_F} ) of the front center wheel side linear damping coefficient C _{s_F} and the front center wheel side variable damping coefficient C _{v_F} . The front center wheel side linear damping coefficient C _{s_F} represents a damping coefficient that does not vary, and the front center wheel side variable damping coefficient C _{v_F} represents a varying damping coefficient. Each linear damping coefficient C _{s_F} , C _{s_rr} , C _{s_rl} is determined in advance. Front center wheel linear damping coefficient C _{S_F,} for example right front wheel suspension system damper predetermined be left wheel side linear damping for predetermined is a right front wheel side linear damping coefficient C _{S_fr} the left front wheel suspension system 10FL of the damper for the 10FR It can be represented by the average of the coefficient C _{s_fl} . In addition, a rear wheel side torsional elastic coefficient that is a coefficient of torsional force generated by the rear wheel side stabilizer 15 is represented by K _{stb_R} .

In S210, the front wheel approximate control unit 212 applies the nonlinear _H∞ control theory to the control system designed based on the equation of motion derived from the front wheel approximate three-wheel model, and the three represented in the model The variable damping coefficients (front center wheel side variable damping coefficient C _{v — F} , right rear wheel side variable damping coefficient C _{v — rr} , left rear wheel side variable damping coefficient C _{v — rl} ) of the suspension devices 10F, 10RR, and 10RL are calculated.

The equation of motion of the sprung member HA derived from the front wheel approximate three-wheel model is expressed by the following equations (eq.26) to (eq.28).

Equation (eq.26) is the heave equation of motion of the sprung member HA, equation (eq.27) is the roll equation of motion, and equation (eq.28) is the pitch equation of motion. In the equation (eq.27), _Tr is a tread on the rear wheel side. In each equation, F _{sus_F} is the force acting on the sprung member HA in the vertical direction at the front center wheel position of the sprung member HA (front center wheel side vertical force), and F _{sus_rr} is the right rear of the sprung member HA. The force acting vertically on the sprung member HA at the wheel position (right rear wheel side vertical force), _{Fsus_rl} is the force acting vertically on the sprung member HA at the left rear wheel position of the sprung member HA ( Left rear wheel side vertical force). FIG. 13 is a diagram showing the action position and acceleration of the force expressed by the equations (eq.26) to (eq.28).

Front center wheel side vertical force F _{sus_F} is expressed by the following equation (eq.29), right rear wheel side vertical force F _{sus_fr} is expressed by the following equation (eq.30), and left rear wheel side vertical force F _{sus_rl} is _expressed by the following equation (eq. 31).

In the above equations (eq.29) to (eq.31), F _{sp_F} is the elastic force (front central wheel side elastic force) generated by the spring of the front center wheel side suspension device 10F, and F _{sp_rr} is the right rear wheel side suspension. Elastic force (right rear wheel side elastic force) generated by the spring of the device 10RR, and F _{sp_rl} is an elastic force (left rear wheel side elastic force) generated by the spring of the left rear wheel side suspension device 10RL. F _{Cs_F} is a damping force (front center wheel side linear damping force) represented by a front center wheel side linear damping coefficient C _{s_F} among damping forces generated by the damper of the front center wheel side suspension device 10F, and F _{Cv_F} is a front side. This is the damping force (front center wheel side variable damping force) represented by the center wheel side variable damping coefficient _{Cv_F} . F _{Cs_rr} is the damping force (right rear wheel side linear damping force) represented by the right rear wheel side linear damping coefficient C _{s_rr} of the damping force generated by the damper of the right rear wheel side suspension device 10RR, and F _{Cv_rr} is the right side. This is the damping force (right rear wheel side variable damping force) represented by the rear wheel side variable damping coefficient _{Cv_rr} . F _{Cs_rl} is the damping force (left rear wheel side linear damping force) represented by the left rear wheel side linear damping coefficient C _{s_rl} of the damping force generated by the damper of the left rear wheel side suspension device 10RL, and F _{Cv_rl} is the left side. This is the damping force (left rear wheel side variable damping force) represented by the rear wheel side variable damping coefficient _{Cv_rl} . F _{stb_R} is a torsional elastic force (rear wheel side torsional elastic force) generated by the rear wheel side stabilizer 15.

The front center wheel side elastic force F _{sp_F} , the right rear wheel side elastic force F _{sp_rr} and the left rear wheel side elastic force F _{sp_rl} are expressed by the following equations (eq.32) to (eq.34).

The front center wheel side linear damping force F _{Cs_F} , the right rear wheel side linear damping force F _{Cs_rr} and the left rear wheel side linear damping force F _{Cs_rl} are expressed by the following equations (eq.35) to (eq.37).

The front center wheel side variable damping force F _{Cv_F} , the right rear wheel side variable damping force F _{Cv_rr} and the left rear wheel side variable damping force F _{Cv_rl} are expressed by the following equations (eq.38) to (eq.40).

The rear wheel side torsional elastic force F _{stb_R} is expressed by the following equation (eq.41).

Further, the heave displacement x _h , the roll angle θ _r and the pitch angle θ _p are calculated by using the mode conversion matrix, as shown in the following equation (eq.42), the front center wheel side sprung displacement x _{b_F} , right The rear wheel side sprung displacement amount x _{b_rr} can be converted into the left rear wheel side sprung displacement amount x _{b_rl} .

From the above relational expression, the state space expression of the front wheel approximate three-wheel model is expressed as the following expression (eq.43).

In the above equation (eq. 43), the state quantity x _p , the evaluation output z _p , and the disturbance w are expressed as the following equation (eq. 44), for example.

The control input u is represented by the front center wheel side variable damping coefficient C _{v_F} , the right rear wheel side variable damping coefficient C _{v_rr} and the left rear wheel side variable damping coefficient C _{v_rl} as shown in the following equation (eq. 45). is there.

Table 3 shows the number of manipulated variables (control input u) in the state space expression shown in Equation (eq.43), the number of state quantities, and the degree of freedom of the front wheel approximate three-wheel model that is the basis of the state space expression. Show.

As shown in Table 3, the degree of freedom of the sprung member represented by the front wheel approximate three-wheel model is three degrees of freedom (vertical motion, roll motion, pitch motion). Further, in this model, the movement of the unsprung member is not considered. That is, the unsprung member is displaced up and down like the road surface. For this reason, the degree of freedom of the suspension device is one degree of freedom (up-and-down motion of the sprung member). Therefore, this model has 3 degrees of freedom. Further, the disturbance of the system is road surface displacement amounts _{xr_F} , _{xr_rr} , and _{xr_rl} of three wheels (front center wheel, right rear wheel, and left rear wheel) as shown in the equation (eq.44). When this disturbance is mode-converted into the displacement of the sprung member HA, the disturbance is expressed by the vertical displacement, roll displacement, and pitch displacement of the sprung member HA. Therefore, the degree of freedom of the sprung member HA due to disturbance is 3 degrees of freedom. Since the degree of freedom of the model and the degree of freedom of the sprung member HA due to disturbance are equal, this control system is controllable.

The front wheel approximate control unit 212 obtains the control input u based on the nonlinear _H∞ control theory using the state space expression, the state quantity, the evaluation output, and the control quantity expressed as the above equation. The details of the analysis method are the same as the method performed using the rear wheel approximate three-wheel model, and thus description thereof is omitted. By obtaining the control input u, the variable damping coefficients C _{v_F} , C _{v_rr} , and C _{v_rl} are obtained.

After calculating the variable damping coefficient, the front wheel approximation control unit 212 proceeds to S212 in FIG. 6, and the second front center wheel side required damping force _{F2req_F} , the second right rear wheel side required damping force _{F2req_rr} , and the second left rear wheel. Calculate the required side damping force F2 _{req_rl} . These required damping forces are damping forces (required damping forces) of the control target that should be generated by the dampers of the respective suspension devices represented in the front wheel approximate three-wheel model. The second front center wheel side required damping force F2 _{req_F} is a required damping force for the front center wheel side suspension device 10F represented by the front wheel approximate three-wheel model, and the front center wheel side linear damping coefficient C _{s_F} and the front center wheel. It is calculated by multiplying the sum (C _{s_F} + C _{v_F} ) with the side variable damping coefficient C _{v_F} by the front center wheel side sprung-road relative speed (x _{r_F} '-x _{b_F} '). The second right rear wheel side required damping force F2 _{req_rr} is a required damping force for the right rear wheel side suspension device 10RR represented in the front wheel approximate three-wheel model, and the right rear wheel side linear damping coefficient C _{s_rr} and the right rear wheel. This is calculated by multiplying the sum (C _{s_rr} + C _{v_rr} ) with the side variable damping coefficient C _{v_rr} by the right rear wheel sprung-road relative speed (x _{r_rr} '-x _{b_rr} '). The second left rear wheel side required damping force F2 _{req_rl} is a required damping force for the left rear wheel side suspension device 10RL represented in the front wheel approximate three-wheel model, and the left rear wheel side linear damping coefficient C _{s_rl} and the left rear wheel. It is calculated by multiplying the sum (C _{s_rl} + C _{v_rl} ) with the side variable damping coefficient C _{v_rl} by the left rear wheel side sprung-road relative speed (x _{r_rl} '-x _{b_rl} '). The processing for calculating the required damping force for the three suspension devices 10F, 10RR, 10RL represented in the front wheel approximate three-wheel model in S212 corresponds to the model required damping force calculation means of the present invention.

Next, in S214, the front wheel approximate control unit 212 _distributes the second front center wheel side required damping force F2 _{req_F} to the second right front wheel side required damping force F2 _{req_fr} and the second left front wheel side required damping force F2 _{req_fl} . . Specifically, by dividing the second front center wheel side required damping force F2 _{req_F} by 2, the second right front wheel side required damping force F2 _{req_fr} and the second left front wheel side required damping force F2 _{req_fl} are calculated. The second right front wheel side required damping force F2 _{req_fr} is a required damping force for the right front wheel side suspension device 10FR calculated based on the front wheel approximate three-wheel model, and the second left front wheel side required damping force F2 _{req_fl} is the left front wheel. This is the required damping force for the side suspension device 10FL. When the mounting position of the front center wheel side suspension device 10F is located between the mounting position of the right front wheel side suspension device 10FR and the mounting position of the left front wheel side suspension device 10FL as in this embodiment, the second right front wheel side The second front center wheel side required damping force F2 _{req_F} so that the required damping force F2 _{req_fr} and the second left front wheel side required damping force F2 _{req_fl} are equal, that is, the distribution ratio is 1: 1. Is divided into the second right front wheel side required damping force F2 _{req_fr} and the second left front wheel side required damping force F2 _{req_fl} . The processing performed in S214 corresponds to the required damping force distribution means of the present invention.

After calculating the second right front wheel side required damping force F2 _{req_fr} and the second left front wheel side required damping force F2 _{req_fl in S214} , the front wheel approximation control unit 212 proceeds to S216, and the second right front wheel side required damping force F2 _{req_fr} , The second left front wheel side required damping force _{F2req_fl} , the second right rear wheel side required damping force _{F2req_rr} , and the second left rear wheel side required damping force _{F2req_rl} are output. Then, it progresses to S218 and complete | finishes this program. As can be seen from the above description, the front wheel approximate control unit 212 of the present embodiment is a control system designed based on the vertical motion, roll motion and pitch motion of the sprung member HA derived from the front wheel approximate three-wheel model. By applying the nonlinear H _∞ control theory, the required damping force for the three suspension devices represented in the front wheel approximate three-wheel model is calculated. Further, of the calculated requested damping force, the requested front damping force for the front center wheel side suspension device 10F (second front center wheel side requested damping force _{F2req_F} ) is replaced by the front center wheel side suspension device 10F. This is distributed to the required damping force (second right front wheel side required damping force F2 _{req_fr} and second left front wheel side required damping force F2 _{req_fl} ) for the side suspension device 10FR and the left front wheel side suspension device 10FL. As a result, the required damping force for the four suspension devices attached to each wheel position of the sprung member HA is calculated. The front wheel approximation control unit 212 corresponds to the second required damping force calculation means of the present invention.

  FIG. 7 is a flowchart showing the flow of the turning state determination program executed by the turning state determination unit 213. The turning state determination unit 213 starts this program in S300 of FIG. Next, in S302, the vehicle speed sensor 226 inputs the vehicle speed V, the steering angle sensor 227 inputs the steered wheel turning angle φ, and the yaw rate sensor 228 inputs the actual yaw rate YR.

  Next, in S304, the target yaw rate YR * estimated from the current running state of the vehicle is calculated. The target yaw rate YR * is an estimated value of the yaw rate derived from the running state of the vehicle, and is calculated based on the vehicle speed V, the turning angle φ, and various values (wheelbase, etc.) representing the reference vehicle model. The process of calculating the target yaw rate in S304 corresponds to the target yaw rate calculating means of the present invention.

  Subsequently, the turning state determination unit 213 determines in S306 that the product of the actual yaw rate YR and the yaw rate deviation (YR * −YR) that is the difference between the target yaw rate YR * and the actual yaw rate YR is positive (or 0). It is determined whether or not there is. When the determination result is Yes (S306: Yes), the process proceeds to S308, and the turning state flag RS is set to US. On the other hand, when the determination result is No (S306: No), the process proceeds to S310, and the turning state flag RS is set to the OS. When the turning state flag RS is set to US, it represents that the turning state of the vehicle is an understeer state (or a neutral steer state). When the turning state flag RS is set to OS, it represents that the turning state of the vehicle is an oversteer state.

  FIG. 14 is a graph showing the relationship between the actual yaw rate YR and the target yaw rate YR * and the turning state of the vehicle. The horizontal axis of the graph is the actual yaw rate YR, and the vertical axis is the yaw rate deviation (YR * -YR). Further, the yaw rate is detected and calculated so as to be positive when turning right and negative when turning left, for example. As can be seen from the figure, when the product YR (YR * -YR) of the actual yaw rate YR and the yaw rate deviation (YR * -YR) is positive, that is, the relationship between the actual yaw rate YR and the yaw rate deviation (YR * -YR). Is represented in the first quadrant or the third quadrant of the graph, the turning state of the vehicle is an understeer state. When the product YR (YR * −YR) is negative, that is, when the relationship between the actual yaw rate YR and the yaw rate deviation (YR * −YR) is expressed in the second quadrant or the fourth quadrant of the graph, The turning state of the vehicle is an oversteer state.

  When the absolute value of the actual yaw rate YR is smaller than the absolute value of the target yaw rate YR *, the relationship between the two is expressed in the first quadrant or the third quadrant of the graph. In this case, since the actual yaw rate YR is smaller than the target yaw rate YR *, the turning radius of the vehicle gradually increases when the vehicle speed is increased at a constant steering angle. Therefore, the turning state of the vehicle becomes an understeer state. When the absolute value of the actual yaw rate YR is larger than the absolute value of the target yaw rate YR *, the relationship between the two is expressed in the second quadrant or the fourth quadrant of the graph. In this case, since the actual yaw rate YR is larger than the target yaw rate YR *, the turning radius of the vehicle gradually decreases when the vehicle speed is increased at a constant steering angle. Therefore, the turning state of the vehicle becomes an oversteer state.

  From the above, it can be determined whether the turning state of the vehicle is an understeer state or an oversteer state based on the sign of the product of the actual yaw rate YR and the yaw rate deviation (YR * −YR). The process of determining whether the product of the actual yaw rate YR and the yaw rate deviation (YR * −YR) is positive or negative in S306 corresponds to the turning state determination means of the present invention.

  In S308 or S310, after setting the turning state flag RS to either US or OS, the turning state determination unit proceeds to S312 and outputs the set turning state flag RS. Then, it progresses to S314 and complete | finishes this program.

FIG. 8 is a flowchart showing the flow of the integrated control program executed by the integrated control unit 214. The integrated control unit 214 starts this program at S400 in the figure. Next, in S402, the first right front wheel side required damping force F1 _{req_fr} , the first left front wheel side required damping force F1 _{req_fl} , the first right rear wheel side required damping force F1 _{req_rr} , the first right wheel approximate control unit 211 calculated in step S402. 1 Input the left rear wheel side required damping force F1 _{req_rl} . Next, in S404, the second right front wheel requested damping force _{F2req_fr} , the second left front wheel requested damping force _{F2req_fl} , the second right rear wheel requested damping force _{F2req_rr} , 2 Input the left rear wheel side required damping force F2 _{req_rl} .

  Next, the integrated control unit 214 inputs the turning state flag RS from the turning state determination unit 213 in S406. Thereafter, the process proceeds to S408, and it is determined whether or not the turning state flag RS is set to US, that is, whether or not the turning state of the vehicle is an understeer state (or a neutral steer state). When this determination result is Yes, the process proceeds to S410. On the other hand, if this determination result is No, that is, if the turning state of the vehicle is an oversteer state, the process proceeds to S412.

In S410, the integrated control unit 214 determines the damping force that should be generated by the dampers of the suspension devices 10FR, 10FL, 10RR, and 10RL based on the requested damping force input from the rear wheel approximate control unit 211. That is, the right front wheel side required damping force _{Freq_fr} representing the damping force to be generated by the damper of the right front wheel side suspension device 10FR is the first right front wheel side required damping force F1 _{req_fr} , and the damper of the left front wheel side suspension device 10FL is generated. The left front wheel requested damping force F _{req_fl} representing the power damping force is the first left front wheel requested damping force F1 _{req_fl} and the right rear wheel requested damping force representing the damping force to be generated by the damper of the right rear wheel suspension device 10RR. F _{req_rr} is the first right rear wheel side required damping force F1 _{Freq_rr} , and the left rear wheel side required damping force F _{req_rl} representing the damping force to be generated by the damper of the left rear wheel side suspension device 10RL is the first left rear wheel side required force. Damping force F1 _{req_rl} is determined respectively.

On the other hand, in S412, the integrated control unit 214 determines the damping force that should be generated by the dampers of the suspension devices 10FR, 10FL, 10RR, and 10RL based on the requested damping force input from the front wheel approximate control unit 212. That is, the right front wheel requested damping force F _{req_fr} is the second right front wheel requested damping force F2 _{req_fr} , the left front wheel requested damping force F _{req_fl} is the second left front wheel requested damping force F2 _{req_fl} , and the right rear wheel requested damping force The force F _{req_rr} is determined as the second right rear wheel requested damping force F2 _{req_rr} , and the left rear wheel requested damping force F _{req_rl} is determined as the second left rear wheel requested damping force F2 _{req_rl} .

After determining the required damping forces F _{req_fr} , F _{req_fl} , F _reqrr , F _{req_rl} to be generated by the dampers of the suspension devices 10FR, 10FL, 10RR, 10RL in S410 or S412, the integrated control unit proceeds to S414 and determines The drive signals representing the requested damping forces F _{req_fr} , F _{req_fl} , F _reqrr , and F _{req_rl} are output to each drive circuit. Then, it progresses to S414 and complete | finishes this program.

  The output drive signals representing the required damping force are input to the corresponding drive circuits 23FR, 23FL, 23RR, and 23RL. Each drive circuit outputs a drive current to the corresponding actuator 132FR, 132FL, 132RR, 132RL based on the input drive signal. Each actuator is driven by this drive current. The corresponding valve 131 is operated by driving the actuator. By operating the valve 131, the damping coefficient (damping force characteristic) of the damper 12 of the suspension device 10 is changed and controlled to a desired magnitude. As a result, the damping force of the four suspension devices attached to each wheel position of the sprung member HA is controlled.

In this way, the integrated control unit 214 is attached to each wheel position of the sprung member HA based on the required damping force calculated by either the rear wheel approximate control unit 211 or the front wheel approximate control unit 212. The damping force for the four suspension devices is controlled. Therefore, the integrated control unit 214, in particular, the processing of S414 for outputting the determined required damping forces _{Freq_fr} , _{Freq_fl} , _Freqrr , _{Freq_rl} to each drive circuit corresponds to the damping force control means of the present invention. Further, the integrated control unit 214 calculates the required damping force calculated using the rear wheel approximate three-wheel model and the front wheel approximate three-wheel model based on the vehicle turning state determined by the turning state determination unit 213. One of the requested damping forces is selected. Specifically, when it is determined that the turning state of the vehicle is an understeer state (S408: Yes), the requested damping force calculated based on the rear wheel approximate three-wheel model is selected (S410), and the process proceeds to S408. When it is determined that the turning state of the vehicle is the oversteer state (S408: No), the requested damping force calculated based on the front wheel approximate three-wheel model is selected (S412). These processes correspond to the required damping force selection means of the present invention.

According to the present embodiment, a three-wheel model (rear wheel approximate three-wheel model or front wheel approximate three-wheel model) is used as the vehicle model, and the control system is designed based on the equation of motion derived from this model. The system becomes controllable. Therefore, the variable damping coefficient can be calculated by applying the non-linear _H∞ control theory to such a controllable control system. Based on the variable damping coefficient calculated in this way, the damping force of the four suspension devices attached to each wheel position of the sprung member HA can be controlled.

Further, the system can be controlled even if the degree of freedom of each suspension device is one degree of freedom considering only the motion of the sprung member. For this reason, the damping force of each suspension device can be controlled without considering the movement of the unsprung member. Therefore, the calculation load on the suspension ECU 21 when the variable damping coefficient is calculated based on the nonlinear _H∞ control theory is reduced.

  Further, when the turning state of the vehicle is an understeer state, the damping force generated by each suspension device is controlled by the required damping force calculated based on the rear wheel approximate three-wheel model. By such damping force control, the suspension devices 10FR and 10FL on the left and right front wheels side generate appropriate damping force according to the turning state so as to suppress the rolling motion of the sprung member HA that occurs as the vehicle turns. . As a result, the side slip on the front wheel side is suppressed, and the turning state of the vehicle approaches the neutral steer state from the under steer state. That is, the understeer state is corrected.

  Further, when the turning state of the vehicle is an oversteer state, the damping force generated in each suspension device is controlled by the required damping force calculated based on the front wheel approximate three-wheel model. By such damping force control, the left and right rear wheel suspension devices 10RR, 10RL generate appropriate damping force according to the turning state so as to suppress the rolling motion of the sprung member HA that occurs as the vehicle turns. To do. As a result, the occurrence of side slip on the rear wheel side is suppressed, and the turning state of the vehicle approaches the neutral steer state from the over steer state. That is, the oversteer state is corrected.

  Thus, when the turning state of the vehicle is an understeer state, the damping force of the suspension device attached to each wheel position of the sprung member is controlled by the required damping force calculated using the rear wheel approximate three-wheel model, When the turning state of the vehicle is an oversteer state, an understeer state is achieved by controlling the damping force of the suspension device attached to each wheel position of the sprung member by the required damping force calculated using the front wheel approximate three-wheel model. Further, since the oversteer state is corrected, the stability during turning is further improved.

(Second Embodiment)
Next, a second embodiment of the present invention will be described. The damping force control apparatus of the first embodiment determines whether to control the damping force using the rear wheel approximate three-wheel model or the front wheel approximate three-wheel model based on the turning state of the vehicle. The damping force control apparatus according to the present embodiment controls the damping force using the rear wheel approximate three-wheel model or uses the front wheel approximate three-wheel model based on whether the sprung acceleration sensor is normal or abnormal. Decide whether to control the damping force. The other parts are the same as those in the first embodiment. Hereinafter, the characteristic part of the present embodiment will be mainly described.

  The overall configuration of the suspension control device according to the present embodiment is the same as that of FIG. 1 described in the first embodiment, and a description thereof will be omitted. FIG. 15 is a diagram illustrating a 20 schematic configuration of the electric control device according to the present embodiment. As shown in the figure, the electric control device 20 according to the present embodiment includes various sensors, a suspension ECU 21, and drive circuits, similarly to the electric control device according to the first embodiment. However, unlike the first embodiment, a vehicle speed sensor, a steering angle sensor, and a yaw rate sensor are not provided. 15, the same components as those of the electric control device 20 (FIG. 3) described in the first embodiment are denoted by the same reference numerals.

  As can be seen from FIG. 15, the electric control device 20 of this embodiment also includes sprung acceleration sensors 221FR, 221FL, 221RR, 221RL. The right front wheel side sprung acceleration sensor 221FR is attached to the right front wheel position of the sprung member HA and detects the vertical acceleration of the sprung member HA at the right front wheel position. The left front wheel side sprung acceleration sensor 221FL is attached to the position of the left front wheel of the sprung member HA, and detects the vertical acceleration of the sprung member HA at the position of the left front wheel. The right rear wheel side sprung acceleration sensor 221RR is attached to the right rear wheel position of the sprung member HA, and detects the vertical acceleration of the sprung member HA at the right rear wheel position. The left rear wheel side sprung acceleration sensor 221RL is attached to the left rear wheel position of the sprung member HA, and detects the vertical acceleration of the sprung member HA at the left rear wheel position. Since other configurations are the same as those in the first embodiment, description thereof is omitted.

FIG. 16 is a diagram showing the suspension ECU 21 for each function. As can be seen from the figure, the suspension ECU 21 includes a rear wheel approximation control unit 211, a front wheel approximation control unit 212, and an integrated control unit 214. The rear wheel approximate control unit 211 is designed based on the heave motion equation, roll motion equation, and pitch motion equation of the sprung member HA derived from the rear wheel approximate three-wheel model, as in the first embodiment. By applying the nonlinear _H∞ control theory to the control system (generalized plant), the damping force to be generated by the dampers of the right front wheel side suspension device 10FR, the left front wheel side suspension device 10FL, and the rear center wheel side suspension device 10R ( First right front wheel side required damping force F1 _{req_fr} , first left front wheel side required damping force F1 _{req_fl} , and first rear center wheel side required damping force F1 _{req_R} ) are respectively calculated. Further, of the calculated required damping force, the required damping force (first rear central wheel side required damping force F1 _{req_R} ) for the rear central wheel side suspension device 10R is the right rear wheel side suspension device 10RR and the left rear wheel side suspension device. The required damping force for 10RL (first right rear wheel side required damping force _{F1req_rr} , first left rear wheel side required damping force _{F1req_rl} ) is distributed. Thus, based on the rear wheel approximate three-wheel model, the first right front wheel side required damping force F1 _{req_fr} , the first left front wheel side required damping force F1 _{req_fl} , the first right rear wheel side required damping force F1 _{req_rr} , the first left The rear wheel side required damping force F1 _{req_rl} is calculated.

Also in the front wheel approximate control unit 212, as in the first embodiment, a control system designed based on the heave motion equation, roll motion equation, and pitch motion equation of the sprung member HA derived from the front wheel approximate three-wheel model. By applying the non-linear _H∞ control theory to the (generalized plant), the damping force to be generated by the dampers of the front center wheel side suspension device 10F, the right rear wheel side suspension device 10RR, and the left rear wheel side suspension device 10RL ( The second front center wheel side required damping force _{F2req_F} , the second right rear wheel side required damping force _{F2req_rr} , and the second left rear wheel side required damping force _{F2req_rl} ) are respectively calculated. Further, of the calculated required damping forces, the required damping force for the front center wheel side suspension device 10F (second front center wheel side required damping force _{F2req_F} ) is obtained for the right front wheel side suspension device 10FR and the left front wheel side suspension device 10FL. Of the required damping force (second right front wheel side required damping force _{F2req_fr} , second left front wheel side required damping force _{F2req_fl} ). Accordingly, based on the front wheel approximate three-wheel model, the second right front wheel side required damping force F2 _{req_fr} , the second left front wheel side required damping force F2 _{req_fl} , the second right rear wheel required damping force F2 _{req_rr} , the second left rear The wheel side required damping force F2 _{req_rl} is calculated respectively.

  In the integrated control unit 214, the required damping force calculated by the rear wheel approximation control unit 211 and the front wheel approximation based on the abnormal state of the sprung acceleration sensor, specifically, which sprung acceleration sensor is abnormal. One of the required damping forces calculated by the control unit 212 is selected, and the damping forces of the four suspension devices attached to the respective wheel positions of the sprung member HA are controlled based on the selected required damping force. . FIG. 17 is a flowchart showing the flow of processing executed by the integrated control unit 214 of this embodiment.

As shown in FIG. 17, the integrated control unit 214 starts the integrated control process in S500 of FIG. Next, in S502, the required damping force calculated by the rear wheel approximation control unit 211 (first right front wheel side required damping force F1 _{req_fr} , first left front wheel side required damping force F1 _{req_fl} , first right rear wheel side required damping force) Force F1 _{req_rr} , first left rear wheel side required damping force F1 _{req_rl} ). Next, in S504, the required damping force calculated by the front wheel approximate control unit 212 (second right front wheel side required damping force F2 _{req_fr} , second left front wheel side required damping force F2 _{req_fl} , second right rear wheel side required damping) Force F2 _{req_rr} , second left rear wheel side required damping force F2 _{req_rl} ).

  Subsequently, in S506, the integrated control unit 214, the right front wheel side sprung acceleration sensor 221FR, the left front wheel side sprung acceleration sensor 221FL, the right rear wheel side sprung acceleration sensor 221RR, and the left rear wheel side sprung acceleration sensor 221RL. Are all operating normally. Whether each sensor is normal or abnormal can be determined by a change in a signal input from each sensor or a comparison between signals input from each sensor.

In this embodiment, when all the sprung acceleration sensors are operating normally, the damping force of each suspension device is controlled by the required damping force calculated based on the rear wheel approximate three-wheel model. Accordingly, if the determination result in S506 is Yes, the process proceeds to S512, in which the right front wheel side required damping force _{Freq_fr} is _set to the first right front wheel side required damping force F1 _{req_fr} , and the left front wheel side required damping force _{Freq_fl} is _set to the first left front wheel. Side required damping force F1 _{req_fl} , right rear wheel side required damping force F _{req_rr} as first right rear wheel side required damping force F1 _{req_rr} , and left rear wheel side required damping force F _{req_rl} as first left rear wheel side required damping force Set each in F1 _{req_rl} . Next, the integrated control unit proceeds to S516, and outputs the set required damping force. Then, it progresses to S520 and complete | finishes this program.

  On the other hand, if the determination result in S506 is No, that is, if at least one of the four sprung acceleration sensors is abnormal, the process proceeds to S508. In S508, the integrated control unit 214 determines whether or not the number of abnormal sprung acceleration sensors is one. If the determination result is Yes, the process proceeds to S510. In S510, the integrated control unit 214 determines whether the sprung acceleration sensor determined to be abnormal is one of the right rear wheel side sprung acceleration sensor 221RR and the left rear wheel side sprung acceleration sensor 221RL. When this determination result is Yes, the process proceeds to S512, and the required damping force calculated based on the rear wheel approximate three-wheel model is selected. Then, the required damping force selected in S516 is output. Thereafter, the process proceeds to S520 and the program is terminated. Thus, in the present embodiment, when the abnormality sensor is one of the four sprung acceleration sensors and the abnormality sensor is a sprung acceleration sensor attached to the rear wheel position of the sprung member HA. The damping force of each suspension device is controlled based on the required damping force calculated using the rear wheel approximate three-wheel model.

If the determination result in S510 is No, that is, if the sprung acceleration sensor determined to be abnormal is one of the right front wheel side sprung acceleration sensor 221FR or the left front wheel side sprung acceleration sensor 221FL, the process proceeds to S514. move on. In S514, the required damping force calculated based on the front wheel approximate three-wheel model is selected. That is, the right front wheel requested damping force F _{req_fr} is the second right front wheel requested damping force F2 _{req_fr} , the left front wheel requested damping force F _{req_fl} is the second left front wheel requested damping force F2 _{req_fl} , and the right rear wheel requested damping force The force F _{req_rr} is _set to the second right rear wheel requested damping force F2 _{req_rr} , and the left rear wheel requested damping force F _{req_rl} is _set to the second left rear wheel requested damping force F2 _{req_rl} . Next, the integrated control unit 214 proceeds to S516, and outputs the set required damping force. Then, it progresses to S520 and complete | finishes this program. As described above, in the present embodiment, when the abnormality sensor is one of the four sprung acceleration sensors and the abnormality sensor is a sprung acceleration sensor attached to the front wheel position of the sprung member HA, Based on the required damping force calculated using the front wheel approximate three-wheel model, the damping force of each suspension device is controlled.

  If the determination result in S508 is No, that is, if there are a plurality of abnormal sensors, the vibration state of the sprung member HA cannot be accurately grasped. In this case, since it is difficult to effectively suppress the vibration of the sprung member HA, the process proceeds to S518 and the variable control of the damping force is prohibited. Then, it progresses to S520 and complete | finishes this program.

According to this embodiment, when either one of the right front wheel side sprung acceleration sensor 221FR or the left front wheel side sprung acceleration sensor 221FL is abnormal and the other sprung acceleration sensor is normal, the front wheel approximation control unit 212 is used. demanded damping force calculated by the (second right front wheel requested damping force _F2 req_fr, second left front wheel requested damping force _F2 req_rl, second right rear wheel requested damping force _F2 req_rr, after the second left wheel side required damping Based on the force _{F2req_rl} ), the damping force of the suspension device attached to each wheel position of the sprung member HA is controlled. Further, when either one of the right rear wheel side sprung acceleration sensor 221RR or the left rear wheel side sprung acceleration sensor 221RL is abnormal and the other sprung acceleration sensor is normal, the rear wheel approximation control unit 211 calculates. _Demanded damping force (first right front wheel side demand damping force F1 _{req_fr} , first left front wheel side demand damping force F1 _{req_fl} , first right rear wheel side demand damping force F1 _{req_rr} , first left rear wheel side demand damping force F1 The damping force of the suspension device attached to each wheel position of the sprung member HA is controlled based on _{req_rl} ).

Thus, four one of the sprung acceleration sensor but also an abnormal, by calculating the required damping force by using the front wheel approximate three-wheeled model or rear approximate three-wheeled model, the nonlinear H _∞ control theory The damping force control of each suspension device based on the above can be continued. Further, since the required damping force is calculated using a three-wheel model that approximates the two suspension devices on the side where the abnormality sensor is installed to one virtual suspension device, the vibration of the sprung member HA is effectively suppressed. Is done.

(Third embodiment)
Next, a third embodiment of the present invention will be described. The damping force control apparatus according to the first and second embodiments is configured to calculate either the required damping force calculated using the rear wheel approximate three-wheel model or the required damping force calculated using the front wheel approximate three-wheel model. The damping force of each suspension device is controlled based on the selected required damping force. On the other hand, in the present embodiment, the three-wheel model to be used is determined in advance as a rear wheel approximate three-wheel model. The damping force control device controls the damping force of each suspension device based on the required damping force calculated using the rear wheel approximate three-wheel model.

  The overall configuration of the suspension control device according to the present embodiment is the same as that of FIG. 1 described in the first embodiment, and a description thereof will be omitted. FIG. 18 is a schematic diagram showing the electric control device 20 according to the present embodiment. As shown in the drawing, the electric control device 20 according to the present embodiment includes various sensors, a suspension ECU 21, and each drive circuit, similarly to the electric control device according to the first embodiment. However, unlike the first embodiment, a vehicle speed sensor, a steering angle sensor, and a yaw rate sensor are not provided. In FIG. 18, the same components as those of the electric control device 20 (FIG. 3) described in the first embodiment are denoted by the same reference numerals, and detailed description thereof is omitted.

  As can be seen from FIG. 18, the suspension ECU 21 includes a rear wheel approximation control unit 211. The suspension ECU 21 of the present embodiment does not have functions corresponding to the front wheel approximate control unit 212, the turning state determination unit 213, and the integrated control unit 214 that the suspension ECU 21 described in the first embodiment has. May be.

In the suspension control device having such a configuration, when controlling the damping force of each suspension device attached to each wheel position of the sprung member, the rear wheel approximation control unit 211 of the suspension ECU 21 executes a rear wheel approximation control program. To do. FIG. 19 is a flowchart showing the flow of the rear wheel approximation control program executed by the rear wheel approximation control unit 211 in the present embodiment. Of the steps shown in this flowchart, the steps S602 to S610 are the same as the steps S102 to S110 in the flowchart of the rear wheel approximation control program of FIG. 5 described in the first embodiment. Represents. Therefore, the rear wheel approximation control unit 211 sequentially executes the processes shown in S602 to S610, so that the right front wheel side variable damping coefficient C _{v — fr} and the left front wheel side variable damping are set as described in the first embodiment. The coefficient C _{v_fl} and the rear center wheel side variable damping coefficient C _{v_R} are calculated.

Next, in S612, the rear wheel approximation control unit 211 sets the right front wheel side sprung-road surface to the sum of the right front wheel side linear damping coefficient C _{s_fr} and the right front wheel side variable damping coefficient C _{v_fr} (C _{s_fr} + C _{v_fr} ). the sum of the between the relative speed to the right front wheel side demanded damping force F _{Req_fr} by multiplying the _{(x r_fr '-x b_fr')} , the left front wheel linear damping coefficient C _{S_fl} and left front side variable damping coefficient _C v_fl (C s_fl + C _{v_fl} ) is _multiplied by the left front wheel side sprung-to-spring relative speed (x _{r_fl} '-x _{b_fl} '), the left front wheel side required damping force F _{req_fl} , the rear center wheel side linear damping coefficient C _{s_R} and the rear center wheel the sum of the side variable damping coefficient _{C v_R (C s_R + C v_R} ) behind the center wheel side sprung - road between the relative velocity _(x r_R '-x b_R') behind the center wheel required damping force by multiplying the F _{req_r} Are calculated respectively. The steps so far correspond to the model required damping force calculation means of the present invention.

Next, the rear wheel approximate control unit 211 divides the rear center wheel side required damping force F _{req_R} by 2 in _S614 to obtain the right rear wheel side required damping force F _{req_rr} and the left rear wheel side required damping force F _{req_rl} . Calculate (required damping force distribution means). In S616, drive signals representing the right front wheel side required damping force _{Freq_fr} , the left front wheel side required damping force _{Freq_fl} , the right rear wheel side required damping force _{Freq_rr} , and the left rear wheel side required damping force _{Freq_rl} are respectively obtained. Output to the corresponding drive circuits 132FR, 132FL, 132RR, 132RL (required damping force control means). Thus, the damping force of each suspension device is controlled based on the required damping force calculated using the rear wheel approximate three-wheel model. Thereafter, the process proceeds to S618 and the program is terminated.

  Thus, according to the present embodiment, the damping force of each suspension device is controlled based on the required damping force calculated using only the rear wheel approximate three-wheel model. When the damping force is controlled using the rear wheel approximate three-wheel model, the left and right suspension devices arranged on the front wheel side generate a damping force that effectively suppresses roll vibration as described above. Therefore, it is preferable to apply the damping force control using the rear wheel approximate three-wheel model to a vehicle that places importance on the turning stability on the front wheel side. For example, damping force control using a rear wheel approximate three-wheel model can be applied to a front-wheel drive FF vehicle or a vehicle that places importance on the stability of the front seat. In addition, for a vehicle in which two sprung acceleration sensors are attached to the front wheel side of the sprung member and one sprung acceleration sensor is attached to the rear wheel side, a rear wheel approximate three-wheel model is used for each wheel position. The damping force of the attached suspension device can be controlled.

(Fourth embodiment)
Next, a fourth embodiment of the present invention will be described. In the present embodiment, the three-wheel model to be used is determined in advance as a front wheel approximate three-wheel model. The damping force control device controls the damping force of each suspension device based on the required damping force calculated using the front wheel approximate three-wheel model.

  The overall configuration of the suspension control device according to the present embodiment is the same as that of FIG. 1 described in the first embodiment, and a description thereof will be omitted. FIG. 20 is a schematic diagram showing the electric control device 20 according to the present embodiment. As shown in the drawing, the electric control device 20 according to the present embodiment includes various sensors, a suspension ECU 21, and each drive circuit, similarly to the electric control device according to the first embodiment. However, unlike the first embodiment, a vehicle speed sensor, a steering angle sensor, and a yaw rate sensor are not provided. In FIG. 20, the same components as those of the electric control device 20 (FIG. 3) described in the first embodiment are denoted by the same reference numerals, and detailed description thereof is omitted.

  As can be seen from FIG. 20, the suspension ECU 21 includes a front wheel approximation control unit 212. The suspension ECU 21 of the present embodiment has functions corresponding to the rear wheel approximate control unit 211, the turning state determination unit 213, and the integrated control unit 214 that the suspension ECU 21 described in the first embodiment has. It does not have to be.

In the suspension control device having such a configuration, when the damping force of each suspension device attached to each wheel position of the sprung member is controlled, the front wheel approximation control unit 212 of the suspension ECU 21 executes a front wheel approximation control program. FIG. 21 is a flowchart showing the flow of the front wheel approximation control program executed by the front wheel approximation control unit 212 in the present embodiment. Of the steps shown in this flowchart, the steps S702 to S710 are the same as the steps S202 to S210 of FIG. 6 which is a flowchart of the front wheel approximation control program described in the first embodiment. Represents. Therefore, the front wheel approximation control unit 212 sequentially executes the processes shown in S702 to S710, so that the front center wheel side variable damping coefficient C _{v_F} and the right rear wheel side variable can be changed as described in the first embodiment. The damping coefficient C _{v_rr} and the left rear wheel side variable damping coefficient C _{v_rl} are calculated.

Next, in S712, the front wheel approximate control unit 212 _adds the front center wheel side linear damping coefficient C _{s_F} and the front center wheel side variable damping coefficient C _{v_F} (C _{s_F} + C _{v_F} ) to the front center wheel side spring top. _-By multiplying the relative speed between road surfaces (x _{r_F} '-x _{b_F} '), the front center wheel side required damping force F _{req_F} is _calculated as the right rear wheel side linear damping coefficient C _{s_rr} and the right rear wheel side variable damping coefficient C _{v_rr} . By multiplying the sum (C _{s_rr} + C _{v_rr} ) by the right rear wheel sprung-to-spring relative speed (x _{r_rr} '-x _{b_rr} '), the right rear wheel requested damping force F _{req_rr} is Left rear by _multiplying the sum of the coefficient C _{s_rl} and the left rear wheel side variable damping coefficient C _{v_rl} (C _{s_rl} + C _{v_rl} ) by the left rear wheel side sprung-road relative speed (x _{r_rl} '-x _{b_rl} ') The wheel side required damping force _{Freq_rl} is calculated. The steps so far correspond to the model required damping force calculation means of the present invention.

Next, the front wheel approximate control unit 212 _calculates the right front wheel side required damping force F _{req_fr} and the left front wheel side required damping force F _{req_fl} by dividing the front center wheel side required damping force F _{req_F} by 2 in S714 ( Required damping force distribution means). In S716, drive signals representing right front wheel side required damping force _{Freq_fr} , left front wheel side required damping force _{Freq_fl} , right rear wheel side required damping force _{Freq_rr} , and left rear wheel side required damping force _{Freq_rl} are respectively obtained. Output to the corresponding drive circuits 132FR, 132FL, 132RR, 132RL (required damping force control means). Thus, the damping force of each suspension device is controlled based on the required damping force calculated using the front wheel approximate three-wheel model. Thereafter, the process proceeds to S718 and the program is terminated.

  Thus, according to the present embodiment, the damping force of each suspension device is controlled based on the required damping force calculated using only the front wheel approximate three-wheel model. When the damping force is controlled using the front wheel approximate three-wheel model, the left and right suspension devices arranged on the rear wheel side generate damping force that effectively suppresses roll vibration as described above. Therefore, it is preferable to apply the damping force control using the front wheel approximate three-wheel model to a vehicle that places importance on the turning stability on the rear wheel side. For example, the damping force control using the front wheel approximate three-wheel model can be applied to an FR vehicle whose driving wheels are rear wheels, a vehicle such as a luxury vehicle and the like that emphasizes stability on the rear seat side. Also, a vehicle with two sprung acceleration sensors attached to the rear wheel side of the sprung member and one sprung acceleration sensor attached to the front wheel side is attached to each wheel position using a front wheel approximate three-wheel model. The damping force of the suspension device provided can be controlled.

  As mentioned above, although embodiment of this invention was described, this invention should not be limited to the said embodiment. For example, in the first embodiment, the turning state determination unit 213 determines whether the turning state of the vehicle is an understeer state based on the sign of the product of the actual yaw rate YR and the yaw rate deviation (YR * −YR). Although it is determined whether the vehicle is present, the turning state of the vehicle may be determined by other methods. Further, the required damping force calculated using the rear wheel approximate three-wheel model based on the turning state of the vehicle in the first embodiment and based on the abnormal state of the sprung acceleration sensor in the second embodiment. Whether the damping force of each suspension device is controlled or whether the damping force of each suspension device is controlled based on the required damping force calculated using the front wheel approximate three-wheel model. However, depending on other factors, such as user preference, select either the required damping force calculated using the rear wheel approximate three-wheel model or the required damping force calculated using the front wheel approximate three-wheel model. You may be able to do it.

  In the first embodiment, the required damping force is calculated using two three-wheel models, and either of the calculated required damping forces is selected based on the turning state of the vehicle, and the selected required damping force is selected. Based on the above, the damping force of each suspension device is controlled. However, the three-wheel model to be used is first determined based on some condition (for example, the turning state of the vehicle), and the damping force of each suspension device is controlled based on the required damping force calculated using the determined three-wheel model. May be. In this case, since the three-wheel model to be used is determined first, it is not necessary to calculate the required damping force based on the three-wheel model that is not used. Therefore, the calculation load is reduced. Thus, the present invention can be modified without departing from the gist thereof.

DESCRIPTION OF SYMBOLS 1 ... Suspension control apparatus, 10FR ... Right front wheel side suspension apparatus, 10FL ... Left front wheel side suspension apparatus, 10F ... Front center wheel side suspension apparatus (virtual suspension apparatus, front side virtual suspension apparatus), 10RR ... Right rear wheel side suspension apparatus DESCRIPTION OF SYMBOLS 10RL ... Left rear wheel side suspension apparatus, 10R ... Rear center wheel side suspension apparatus (virtual suspension apparatus, rear virtual suspension apparatus), 11 ... Spring, 12 ... Damper, 13 ... Variable throttle mechanism, 131 ... Valve, 132 ... Actuators, 14 ... front wheel side stabilizers, 15 ... rear wheel side stabilizers, 20 ... electric control device (damping force control device), 211 ... rear wheel approximation control unit (first required damping force calculation means), 212 ... front wheel approximation control unit (Second required damping force calculation means), 213... 214, integrated control unit (damping force control means), 221FR ... right front wheel side sprung acceleration sensor, 221FL ... left front wheel side sprung acceleration sensor, 221RR ... right rear wheel side sprung acceleration sensor, 221RL ... left rear wheel Side sprung acceleration sensor, 228... Yaw rate sensor (yaw rate detection sensor), 21... Suspension ECU, HA... Sprung member, LA.

Claims (10)

In a damping force control device for controlling damping force of four suspension devices attached to each wheel position of a sprung member of a vehicle,
The two front wheel suspension devices attached to the left and right front wheel positions of the sprung member or the two rear wheel suspension devices attached to the left and right rear wheel positions of the sprung member were replaced with one virtual suspension device. By applying nonlinear _H∞ control theory to a control system designed based on the vertical motion, roll motion and pitch motion of the sprung member derived from the three-wheel model representing the motion of the vehicle, the three-wheel model is represented by The required damping force, which is the damping force of the control target to be generated by the three suspension devices, is calculated, and the required damping force for the virtual suspension device among the calculated required damping forces is replaced with the virtual suspension device. By distributing the required damping force for the two suspension devices, the position of each sprung member can be adjusted. A required damping force calculating means for calculating the required damping force for the four attached suspension devices;
Damping force control means for controlling the damping force of the four suspension devices attached to each wheel position of the sprung member based on the requested damping force calculated by the requested damping force calculation means;
A damping force control device comprising: The damping force control apparatus according to claim 1,
The required damping force calculation means includes:
Model required damping force calculation means for calculating the required damping force for the three suspension devices represented in the three-wheel model, and the virtual suspension apparatus among the required damping forces calculated by the model required damping force calculation means A damping force control device comprising: a damping force distribution unit that distributes the damping force requested to the damping forces required for the two suspension devices replaced by the virtual suspension device. The damping force control device according to claim 1 or 2,
The three-wheel model is a rear-wheel approximate three-wheel model representing a vehicle motion in which the two rear-wheel suspension devices are replaced with one rear-wheel virtual suspension device, and the damping force is characterized in that Control device. The damping force control device according to claim 1 or 2,
The three-wheel model is a front-wheel approximate three-wheel model representing a motion of a vehicle in which the two front wheel side suspension devices are replaced with one front wheel side virtual suspension device. The damping force control device according to claim 1 or 2,
The required damping force calculation means includes:
Up and down motion, roll motion and pitch motion of a sprung member derived from a rear wheel approximate three-wheel model representing the motion of a vehicle in which the two rear wheel suspension devices are replaced with one rear wheel virtual suspension device. By applying nonlinear _H∞ control theory to a control system designed based on the above, the required damping force for the three suspension devices represented in the rear wheel approximate three-wheel model is calculated, and the calculated required damping force 4 of the sprung member attached to each wheel position by distributing the required damping force for the rear wheel side virtual suspension device to the required damping force for the two rear wheel side suspension devices. A first required damping force calculating means for calculating a required damping force for the suspension device;
Based on the vertical motion, roll motion and pitch motion of the sprung member derived from the front wheel approximate three-wheel model representing the motion of the vehicle in which the two front wheel suspension devices are replaced with one front wheel virtual suspension device. By applying nonlinear _H∞ control theory to the control system to be designed, the required damping force for the three suspension devices represented in the front wheel approximate three-wheel model is calculated, and the front wheel of the calculated required damping force is calculated. The required damping force for the four suspension devices attached to each wheel position of the sprung member is distributed by distributing the required damping force for the side virtual suspension device to the required damping force for the two front wheel side suspension devices. A second required damping force calculating means for calculating force,
The damping force control means is attached to each wheel position of the sprung member based on the demand damping force calculated by one of the first demand damping force calculation means and the second demand damping force calculation means. A damping force control device that controls damping force of each suspension device. In the damping force control device according to claim 5,
The damping force control means is one of the required damping force calculated by the first required damping force calculation means and the required damping force calculated by the second required damping force calculation means based on the turning state of the vehicle. Required damping force selection means for selecting the damping force, and based on the requested damping force selected by the requested damping force selection means, the damping forces of the four suspension devices attached to the respective wheel positions of the sprung member are controlled. A damping force control device. The damping force control apparatus according to claim 6,
The required damping force selection means selects the required damping force calculated by the first required damping force calculation means when the turning state of the vehicle is an understeer state, and the required damping force is selected when the turning state of the vehicle is an oversteer state. (2) A damping force control device, wherein the requested damping force calculated by the requested damping force calculation means is selected. The damping force control device according to claim 6 or 7,
A yaw rate detection sensor for detecting an actual yaw rate generated when the vehicle turns,
A target yaw rate calculating means for calculating a target yaw rate based on the vehicle speed and the steering angle;
Based on the yaw rate deviation representing the difference between the actual yaw rate detected by the yaw rate detection sensor and the target yaw rate calculated by the target yaw rate calculation means, it is determined whether the turning state of the vehicle is an understeer state or an oversteer state. A turning state determining means for determining,
The requested damping force selection unit selects a requested damping force based on the turning state of the vehicle determined by the turning state determination unit. In the damping force control device according to claim 5,
A right front wheel-side sprung acceleration sensor that detects vertical acceleration of the sprung member at the right front wheel position, a left front wheel-side sprung acceleration sensor that detects vertical acceleration of the sprung member at the left front wheel position, and a spring at the right rear wheel position A right rear wheel side sprung acceleration sensor for detecting the vertical acceleration of the upper member, and a left rear wheel side sprung acceleration sensor for detecting the vertical acceleration of the sprung member at the position of the left rear wheel,
When either one of the right front wheel side sprung acceleration sensor or the left front wheel side sprung acceleration sensor is abnormal, the damping force control means sets the required damping force calculated by the second required damping force calculation means. Based on the control, the damping force of the suspension device attached to each wheel position of the sprung member is controlled, and when either one of the right rear wheel side sprung acceleration sensor or the left rear wheel side sprung acceleration sensor is abnormal The damping force control device controls the damping force of the suspension device attached to each wheel position of the sprung member based on the requested damping force calculated by the first required damping force calculation means. The damping force control apparatus according to claim 1,
The three-wheel model is designed so that the virtual suspension device is attached to the sprung member at an intermediate position between the two suspension devices to be replaced are attached to the sprung member. A damping force control device.

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