阅读说明：本技术 车辆的行驶状态控制装置以及车辆的行驶状态控制方法 (Vehicle running state control device and vehicle running state control method ) 是由 古田浩贵 于 2021-06-04 设计创作，主要内容包括：本发明涉及车辆的行驶状态控制装置以及车辆的行驶状态控制方法。控制单元被配置为执行减振控制和侧倾控制,该减振控制是在车轮从车轮通过预测位置通过时基于目标减振控制力来控制控制力产生装置的控制,该侧倾控制是基于目标侧倾控制力来控制控制力产生装置的控制,该目标侧倾控制力用于基于侧倾指标值来减小簧上的侧倾,在同时执行减振控制和侧倾控制时,基于进行了目标减振控制力的减小校正和目标侧倾控制力的增大校正中的至少一方后的目标减振控制力和目标侧倾控制力来控制控制力产生装置。(The present invention relates to a vehicle running state control device and a vehicle running state control method. The control unit is configured to execute vibration damping control that is control of controlling the control-force generating device based on a target vibration damping control force for reducing a roll on the spring based on a roll index value when the wheel passes from the wheel through the predicted position, and roll control that is control of controlling the control-force generating device based on the target roll control force, the control-force generating device being controlled based on the target vibration damping control force and the target roll control force after at least one of a reduction correction of the target vibration damping control force and an increase correction of the target roll control force is performed when the vibration damping control and the roll control are executed simultaneously.)

1. A running state control device for a vehicle, characterized by comprising:

a control force generation device configured to generate a control force in a vertical direction for damping vibration on a spring of a vehicle between at least a pair of right and left wheels and a vehicle body portion corresponding to the positions of the wheels;

a road surface displacement-related information acquisition device configured to acquire road surface displacement-related information associated with an up-down displacement of a road surface;

a roll index value acquisition device configured to acquire a roll index value representing a degree of roll on the spring; and

a control unit configured to control the control-force generating device based on at least one of the road surface displacement-related information and the roll index value,

wherein the control unit is configured to perform: vibration damping control of determining a wheel passage prediction position predicted to be passed by the wheel, calculating a target vibration damping control force for reducing vibration on the spring when the wheel passes from the wheel passage prediction position based on the road surface displacement-related information, and controlling the control force generation device based on the target vibration damping control force when the wheel passes from the wheel passage prediction position; and roll control that calculates a target roll control force for reducing the roll on the spring based on the roll index value, controls the control-force generating device based on the target roll control force,

the control unit is configured to: when the vibration damping control and the roll control are simultaneously executed, the control-force generating device is controlled based on a target vibration damping control force and a target roll control force, at which at least one of a decrease correction of the target vibration damping control force and an increase correction of the target roll control force is performed.

2. The running state control apparatus of a vehicle according to claim 1,

the control unit is configured to: the reduction correction amount of the target damping control force is determined based on the roll index value on the spring such that the larger the magnitude of the roll index value on the spring, the larger the reduction correction amount of the target damping control force.

3. The running state control apparatus of a vehicle according to claim 1 or 2,

the control unit is configured to: the increase correction amount of the target roll control force is determined based on the roll index value on the spring such that the increase correction amount of the target roll control force is larger as the magnitude of the roll index value on the spring is larger.

4. The running state control apparatus of a vehicle according to any one of claims 1 to 3,

the sprung roll index value is any one of an estimated lateral acceleration of the vehicle, an actual lateral acceleration of the vehicle, and a roll angle on the spring.

5. The running state control apparatus of a vehicle according to claim 1 or 2,

the control unit is configured to: an index value of the vibration damping control indicating the magnitude of the control force of the vibration damping control is calculated, and the increase correction amount of the target roll control force is determined based on the index value of the vibration damping control such that the increase correction amount of the target roll control force increases as the index value of the vibration damping control increases.

6. A method of controlling a running state of a vehicle, the method controlling the running state of the vehicle by controlling a control force generating device configured to generate a control force in a vertical direction for damping a vibration on a spring of the vehicle between at least a pair of right and left wheels and a vehicle body portion corresponding to a position of the wheel,

including the vibration damping control and the roll control,

the vibration damping control includes: acquiring road surface displacement associated information associated with the vertical displacement of the road surface; a step of deciding a wheel passing predicted position that the wheel is predicted to pass; a step of calculating a target vibration damping control force for reducing vibration on the spring when the wheel passes through a predicted position from the wheel, based on the road surface displacement-related information; and a step of controlling the control-force generating device based on the target vibration damping control force when the wheel passes from the wheel passing predicted position,

the roll control includes: a step of acquiring a roll index value indicating a degree of roll on the spring; calculating a target roll control force for reducing the roll on the spring based on the roll index value when the roll index value is equal to or greater than a reference value; and a step of controlling the control-force generating device based on the target roll control force,

wherein the control-force generating device is controlled based on a target damping control force and a target rolling control force, which have been subjected to at least one of a decrease correction of the target damping control force and an increase correction of the target rolling control force, when the damping control and the rolling control are executed simultaneously.

Technical Field

The present invention relates to a vehicle running state control device and a vehicle running state control method.

Background

In the control of the vehicle running state, in the control of damping vibration on the spring, a control force in the vertical direction generated between the wheel and the vehicle body is controlled in order to damp vibration on the spring. As sprung vibration damping control of a vehicle, for example, as described in japanese patent laid-open No. 5-319066 described below, predictive vibration damping control is known in which sprung vibration damping is performed at a front wheel position and a rear wheel position using a road surface displacement in the vertical direction in front of the vehicle acquired by a predictive sensor. According to the predictive vibration damping control, the sprung mass can be effectively damped without delay, as compared with the vibration damping control that is performed based on the detection result of the motion state amount in the vertical direction of the vehicle, such as the vertical acceleration of the sprung mass.

As control for damping vibration on the spring of the vehicle, as described in, for example, the following U.S. patent application publication No. 2018/154723, predictive vibration damping control is also known which is performed based on predictive reference data (road surface information acquired in advance) including position information and road surface information of the vehicle. It is foreseen that the reference data is stored on a server with which the vehicle can communicate wirelessly. The road surface information included in the prediction reference data is a value indicating a vertical displacement of the road surface (road surface displacement), and is generated based on sensing data acquired by a prediction sensor such as a camera sensor, a LIDAR (Light Detection and Ranging), a radar, or a plane or three-dimensional scanning sensor.

As control of the running state of the vehicle, roll control is also known in which, when a roll index value indicating the degree of roll on the spring is equal to or greater than a reference value, a target roll control force for reducing the roll on the spring is calculated based on the roll index value, and the control force is controlled based on the target roll control force. According to such roll control, the roll on the spring can be reduced as compared with the case where the control force is not controlled based on the target roll control force, thereby improving the steering stability of the vehicle.

In a vehicle capable of executing both vibration damping control and roll control for controlling a vertical control force generated between a wheel and a vehicle body, there are cases where vibration damping control and roll control are executed simultaneously. Even if roll control is executed, roll moment acts on the spring when the control forces of the left and right vibration damping control are in opposite phases. Therefore, the amount of roll on the spring during turning of the vehicle may increase as compared to the case where the sprung vibration damping control is not performed.

Further, in a situation where the vibration damping control and the roll control are executed simultaneously, the roll index value may change due to a change in the roll angle on the spring and the rate of change thereof by the vibration damping control. In particular, when the roll index value is decreased, the target roll control force may be insufficient, failing to effectively reduce the roll on the spring.

Disclosure of Invention

The invention provides a vehicle running state control device and method capable of reducing roll on a spring more effectively than before under the condition of simultaneously executing vibration damping control and roll control.

According to an aspect of the present invention, there is provided a running state control device for a vehicle, including: a control force generation device configured to generate a control force in a vertical direction for damping vibration on a spring of a vehicle between at least a pair of right and left wheels and a vehicle body portion corresponding to the positions of the wheels; a road surface displacement-related information acquisition device configured to acquire road surface displacement-related information associated with an up-down displacement of a road surface; a roll index value acquisition device configured to acquire a roll index value representing a degree of roll on the spring; and a control unit configured to control the control-force generating device based on at least one of the road-surface-displacement-related information and the roll index value.

The control unit is configured to perform: a vibration damping control that determines a wheel passing predicted position at which the wheel is predicted to pass, calculates a target vibration damping control force for reducing vibration on a spring when the wheel passes from the wheel passing predicted position based on road surface displacement-related information, and controls the control force generation device based on the target vibration damping control force when the wheel passes from the wheel passing predicted position; and roll control for calculating a target roll control force for reducing the roll on the spring based on the roll index value, and controlling the control-force generating device based on the target roll control force.

The control unit is configured to: when the vibration damping control and the roll control are simultaneously executed, the control-force generating device is controlled based on the target vibration damping control force and the target roll control force after at least one of the decrease correction of the target vibration damping control force and the increase correction of the target roll control force is performed.

According to the above configuration, the vibration damping control is performed to control the control force generating device based on the target vibration damping control force for reducing the vibration on the spring when the wheel passes through the predicted position from the wheel, so the vibration on the spring can be reduced. Further, since the roll control is performed by calculating the target roll control force for reducing the roll on the spring based on the roll index value and controlling the control-force generating device based on the target roll control force, the roll on the spring can be reduced.

When the vibration damping control and the roll control are simultaneously executed, the control-force generating device is controlled based on the target vibration damping control force and the target roll control force after at least one of the decrease correction of the target vibration damping control force and the increase correction of the target roll control force is performed.

Thus, in a situation where the vibration damping control and the roll control are executed simultaneously, the possibility of the roll on the spring being deteriorated by the control force of the vibration damping control can be reduced as compared with a case where neither the reduction correction nor the increase correction of the target vibration damping control force is performed.

In the above aspect, it may be that the control unit is configured to: the reduction correction amount of the target damping control force is determined based on the roll index value on the spring so that the larger the magnitude of the roll index value on the spring, the larger the reduction correction amount of the target damping control force.

According to the above aspect, the reduction correction amount of the target damping control force can be changed in accordance with the sprung roll index value such that the larger the magnitude of the sprung roll index value, the larger the reduction correction amount of the target damping control force. Accordingly, the higher the possibility that the roll on the spring becomes larger, the larger the amount by which the control force of the vibration damping control is reduced, and therefore, the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control can be appropriately reduced as compared with the case where the reduction correction amount of the target vibration damping control force is constant.

In the above aspect, it may be that the control unit is configured to: the increase correction amount of the target roll control force is determined based on the roll index value on the spring so that the increase correction amount of the target roll control force increases as the magnitude of the roll index value on the spring increases.

According to the above aspect, the increase correction amount of the target roll control force can be changed in accordance with the roll index value on the spring so that the increase correction amount of the target roll control force becomes larger as the magnitude of the roll index value on the spring becomes larger. Accordingly, the higher the possibility that the roll on the spring becomes larger, the larger the amount by which the effect of the roll control increases, and therefore, the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control can be appropriately reduced as compared with the case where the increase correction amount of the target roll control force is constant.

In the above aspect, the sprung roll index value may be any one of an estimated lateral acceleration of the vehicle, an actual lateral acceleration of the vehicle, and a sprung roll angle.

According to the above aspect, at least one of the reduction correction of the target vibration damping control force and the increase correction of the target roll control force can be performed based on any one of the estimated lateral acceleration of the vehicle, the actual lateral acceleration of the vehicle, and the sprung roll angle.

In the above aspect, it may be that the control unit is configured to: an index value of the vibration damping control indicating the magnitude of the control force of the vibration damping control is calculated, and the increase correction amount of the target roll control force is determined based on the index value of the vibration damping control so that the increase correction amount of the target roll control force increases as the index value of the vibration damping control increases.

According to the above aspect, the increase correction amount of the target roll control force can be changed in accordance with the index value of the vibration damping control so that the increase correction amount of the target roll control force increases as the index value of the vibration damping control indicating the magnitude of the control force of the vibration damping control increases. Accordingly, the higher the possibility that the roll on the spring is deteriorated by the vibration damping control force, the larger the amount by which the effect of the roll control is increased, and therefore, the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control can be appropriately reduced as compared with the case where the increase correction amount of the target roll control force is constant.

According to another aspect of the present invention, there is provided a running state control method of a vehicle for controlling a running state of the vehicle by controlling a control-force generating device configured to generate a control force in a vertical direction for damping a sprung mass of the vehicle between at least a pair of right and left wheels and a vehicle body portion corresponding to a position of the wheels, the method comprising, when simultaneously performing damping control and roll control, controlling the control-force generating device based on a target damping control force and a target roll control force after performing at least one of a decrease correction of the target damping control force and an increase correction of the target roll control force, wherein the damping control comprises: acquiring road surface displacement associated information associated with the vertical displacement of the road surface; a step of deciding a wheel passing predicted position that the wheel is predicted to pass; calculating a target vibration damping control force for reducing vibration on a spring when the wheel passes through the predicted position from the wheel, based on the road surface displacement-related information; and a step of controlling the control-force generating device based on the target damping control force when the wheel passes from the wheel passing predicted position, the roll control including: a step of acquiring a roll index value indicating a degree of roll on the spring; calculating a target roll control force for reducing the roll on the spring based on the roll index value; and a step of controlling the control-force generating device based on the target roll control force.

According to the control method described above, the vibration damping control is performed that controls the control force generation device based on the target vibration damping control force for reducing the vibration on the spring when the wheel passes from the wheel passing through the predicted position, so the vibration on the spring can be reduced. Further, since the roll control is performed by calculating the target roll control force for reducing the roll on the spring based on the roll index value and controlling the control-force generating device based on the target roll control force, the roll on the spring can be reduced.

When the vibration damping control and the roll control are simultaneously executed, the control-force generating device is controlled based on the target vibration damping control force and the target roll control force after at least one of the decrease correction of the target vibration damping control force and the increase correction of the target roll control force is performed. Thus, in a situation where the vibration damping control and the roll control are executed simultaneously, the possibility of the roll on the spring being deteriorated by the control force of the vibration damping control can be reduced as compared with a case where neither the reduction correction nor the increase correction of the target vibration damping control force is performed.

In the present application, the "road surface displacement-related information" may be at least one of an unsprung displacement indicating an unsprung vertical displacement of the vehicle, an unsprung velocity as a time differential value of the unsprung displacement, a road surface displacement indicating a vertical displacement of the road surface, and a road surface displacement velocity as a time differential value of the road surface displacement. Further, the "road surface displacement-related value" may be one of an unsprung displacement representing an unsprung vertical displacement of the vehicle and a road surface displacement representing a vertical displacement of the road surface.

Drawings

Features, advantages and technical and industrial significance of exemplary embodiments of the present invention will be described below with reference to the accompanying drawings, wherein like reference numerals denote like elements, and wherein:

fig. 1 is a schematic configuration diagram of a vehicle to which a running state control device according to an embodiment of the present invention is applied.

Fig. 2 is a schematic configuration diagram of a traveling state control device according to a first embodiment of the present invention.

Fig. 3 is a diagram showing a single-wheel model of the vehicle.

Fig. 4 is a diagram for explaining predictive vibration damping control.

Fig. 5 is another diagram for explaining predictive vibration damping control.

Fig. 6 is another diagram for explaining predictive vibration damping control.

Fig. 7 is a flowchart showing a running state control routine of the vehicle according to the first embodiment.

Fig. 8 is a flowchart showing a subroutine executed in step 750 of fig. 7.

Fig. 9 is a diagram showing a map for calculating the target roll control force Frit based on the estimated lateral acceleration Gyh of the vehicle.

Fig. 10 is a diagram showing a map for calculating the correction coefficient Ac for the target vibration damping control force Fcit based on the absolute value of the estimated lateral acceleration Gyh of the vehicle.

Fig. 11 is a schematic configuration diagram of a traveling state control device according to a second embodiment of the present invention.

Fig. 12 is a flowchart showing a running state control routine of the vehicle according to the second embodiment.

FIG. 13 is a flow chart showing a subroutine executed in step 1250 of FIG. 12.

Fig. 14 is a diagram showing a map for calculating the target roll control force Frit based on the actual lateral acceleration Gy of the vehicle.

Fig. 15 is a diagram showing a map for calculating the correction coefficient Ar for the target roll control force Frit based on the absolute value of the actual lateral acceleration Gy of the vehicle.

Fig. 16 is a flowchart showing a running state control routine of the vehicle according to the third embodiment.

Fig. 17 is a diagram showing a map for calculating a correction coefficient Ar for the target roll control force Frit based on the index value Fca of the vibration damping control.

Detailed Description

First embodiment

Form a

A running state control device of a vehicle of a first embodiment of the present invention is shown in fig. 2 as a whole by reference numeral 20, and this running state control device 20 is applied to the vehicle 10 shown in fig. 1.

The vehicle 10 includes a left front wheel 11FL, a right front wheel 11FR, a left rear wheel 11RL, and a right rear wheel 11 RR. The left front wheel 11FL is rotatably supported by a wheel support member 12 FL. The right front wheel 11FR is rotatably supported by a wheel support member 12 FR. The left rear wheel 11RL is rotatably supported by a wheel supporting member 12 RL. The right rear wheel 11RR is rotatably supported by a wheel support member 12 RR.

The front left wheel 11FL, the front right wheel 11FR, the rear left wheel 11RL, and the rear right wheel 11RR are referred to as "wheels 11" without distinguishing them. Likewise, the left front wheel 11FL and the right front wheel 11FR are referred to as "front wheels 11F". Likewise, the left rear wheel 11RL and the right rear wheel 11RR are referred to as "rear wheels 11R". The wheel supporting members 12FL to 12RR are referred to as "wheel supporting members 12".

The vehicle 10 further includes a front left wheel suspension 13FL, a front right wheel suspension 13FR, a rear left wheel suspension 13RL, and a rear right wheel suspension 13 RR. Details of the suspensions 13FL to 13RR are described below. Note that the suspensions 13FL to 13RR are preferably independently suspended suspensions.

The front left wheel suspension 13FL suspends the front left wheel 11FL from the vehicle body 10a, and includes a suspension arm 14FL, a shock absorber 15FL, and a suspension spring 16 FL. The right front wheel suspension 13FR suspends the right front wheel 11FR from the vehicle body 10a, and includes a suspension arm 14FR, a shock absorber 15FR, and a suspension spring 16 FR.

The left rear wheel suspension 13RL suspends the left rear wheel 11RL from the vehicle body 10a, and includes a suspension arm 14RL, a shock absorber 15RL, and a suspension spring 16 RL. The right rear wheel suspension 13RR suspends the right rear wheel 11RR from the vehicle body 10a, and includes a suspension arm 14RR, a shock absorber 15RR, and a suspension spring 16 RR.

Note that the front left wheel suspension 13FL, the front right wheel suspension 13FR, the rear left wheel suspension 13RL, and the rear right wheel suspension 13RR are referred to as "suspensions 13" without distinguishing them. Likewise, the suspension arms 14FL to 14RR are referred to as "suspension arms 14". Likewise, the dampers 15FL to 15RR are referred to as "dampers 15". Likewise, the suspension springs 16FL to 16RR are referred to as "suspension springs 16".

The suspension arm 14 connects the wheel support member 12 to the vehicle body 10 a. In fig. 1, only one suspension arm 14 is illustrated for one suspension 13, but a plurality of suspension arms 14 may be provided for one suspension 13.

The shock absorber 15 is disposed between the vehicle body 10a and the suspension arm 14, and is coupled to the vehicle body 10a at an upper end and to the suspension arm 14 at a lower end. The suspension spring 16 is elastically mounted between the vehicle body 10a and the suspension arm 14 via a shock absorber 15. That is, the upper end of the suspension spring 16 is connected to the vehicle body 10a, and the lower end of the suspension spring 16 is connected to the cylinder of the shock absorber 15. In the elastic attachment method of the suspension spring 16, the damper 15 may be disposed between the vehicle body 10a and the wheel support member 12.

In this example, the damper 15 is a damping force non-variable type damper, but the damper 15 may be a damping force variable type damper. The suspension spring 16 may be elastically mounted between the vehicle body 10a and the suspension arm 14 without interposing the shock absorber 15 therebetween. That is, the upper end of the suspension spring 16 may be coupled to the vehicle body 10a, and the lower end of the suspension spring 16 may be coupled to the suspension arm 14. In the elastic attachment method of the suspension spring 16, the damper 15 and the suspension spring 16 may be disposed between the vehicle body 10a and the wheel support member 12.

A portion of the member such as the wheel 11 and the damper 15 of the vehicle 10 on the wheel 11 side of the suspension spring 16 is referred to as "unsprung portion 50 (see fig. 3)". In contrast, a portion of the vehicle 10 on the vehicle body 10a side of the suspension spring 16 among the members such as the vehicle body 10a and the shock absorber 15 is referred to as "sprung portion 51 (see fig. 3)".

Further, a front left wheel active actuator 17FL, a front right wheel active actuator 17FR, a rear left wheel active actuator 17RL and a rear right wheel active actuator 17RR are provided between the vehicle body 10a and the suspension arms 14FL to 14RR, respectively. The active actuators 17FL to 17RR are provided side by side with the shock absorbers 15FL to 15RR and the suspension springs 16FL to 16RR, respectively.

Note that the front left wheel active actuator 17FL, the front right wheel active actuator 17FR, the rear left wheel active actuator 17RL, and the rear right wheel active actuator 17RR are referred to as "active actuators 17" without distinguishing them. Likewise, the left front wheel active actuator 17FL and the right front wheel active actuator 17FR are referred to as "front wheel active actuators 17F". Likewise, the left rear wheel active actuator 17RL and the right rear wheel active actuator 17RR are referred to as "rear wheel active actuators 17R".

The active actuator 17 functions as an actuator that variably generates a force (hereinafter, referred to as a "control force") Fc in the vertical direction acting between the vehicle body 10a and the wheel 11 (between the sprung portion 51 and the unsprung portion 50) in order to damp the sprung portion 51, based on a control command from an electronic control device (hereinafter, referred to as an "ECU", and may also be referred to as "control means") 30 shown in fig. 2. The active actuator 17 is also sometimes referred to as a "control-force generating device". In the present example, the active actuator 17 is an electromagnetic active actuator. The active actuator 17 constitutes an active suspension in cooperation with the shock absorber 15, the suspension spring 16, and the like.

In the first embodiment, as shown in fig. 2, the running state control device 20 includes a device 21 mounted on the vehicle and a device 22 mounted outside the vehicle. The in-vehicle device 21 includes the ECU30, the storage device 30a, the positional information acquisition device 33, and the wireless communication device 34. Further, the vehicle-mounted device 21 includes the above-described active actuators 17FL to 17 RR.

The ECU30 includes a microcomputer. The microcomputer includes a CPU, a ROM, a RAM, an interface (I/F), and the like. The CPU realizes various functions by executing instructions (programs, routines) stored in the ROM.

The ECU30 is connected to a nonvolatile storage device 30a capable of reading and writing information. In this example, the storage device 30a is a hard disk drive. The ECU30 can store (save) information in the storage device 30a and read out information stored (saved) in the storage device 30 a. The storage device 30a is not limited to a hard disk drive, and may be a well-known storage device or storage medium that can read and write information.

The vehicle-mounted device 21 further includes a steering angle sensor 31 and a switch 35. The steering angle sensor is a vehicle-mounted sensor and is connected to the ECU 30. The steering angle sensor 31 detects a rotation angle of a steering shaft of a steering device, not shown in the drawings, as a steering operation amount of the driver. The switch 35 is operated by the occupant of the vehicle 10, and the ECU30 performs a predictive vibration damping control described later when the switch 35 is on.

Further, the ECU30 is connected to the position information acquiring device 33 and the wireless communication device 34.

The position information acquiring device 33 includes a GNSS (Global Navigation Satellite System) receiver and a map database. The GNSS receiver receives "signals from artificial satellites (for example, GNSS signals)" for detecting the position of the vehicle 10 at the present time (current position). Road map information and the like are stored in the map database. The position information acquisition device 33 is a device, such as a navigation device, that acquires the current position (e.g., latitude and longitude) of the vehicle 10 based on the GNSS signal.

The ECU30 acquires the vehicle speed V1 of the vehicle 10 and the traveling direction Td of the vehicle 10 at the current time based on the history of the current position acquired by the position information acquiring device 33. Note that, as shown in phantom lines in fig. 2, the vehicle speed V1 may be detected by the vehicle speed sensor 32.

Wireless communication device 34 is a wireless communication terminal for communicating information with cloud 40 of device 22 outside the vehicle via a network. The cloud 40 includes a "management server 42 and a plurality of storage devices 44A to 44N" connected to a network. One or more of storage devices 44A-44N are referred to as "storage device 44" without distinguishing between them. Storage device 44 functions as a storage device outside the vehicle of running state control device 20.

The management server 42 includes a CPU, a ROM, a RAM, an interface (I/F), and the like. The management server 42 performs retrieval and reading of data stored in the storage device 44, and writes data to the storage device 44.

The memory device 44 stores forecast reference data 45. Unsprung displacement z acquired based on the vertical movement state quantity of the vehicle 10 or another vehicle detected when the vehicle 10 or another vehicle actually travels on the road surface 55₁The predicted reference data 45 is registered in association with information on the position where the motion state quantity is detected. Thus, the prediction reference data 45 is the unsprung displacement z acquired based on the vertical movement state quantity of the vehicle 10 or another vehicle₁And data combined with information of the position where the motion state quantity is detected.

The unsprung portion 50 is displaced in the vertical direction by receiving the displacement of the road surface 55 when the vehicle 10 travels on the road surface 55. Unsprung displacement z₁Is the vertical displacement of the unsprung mass 50 corresponding to the position of each wheel 11 of the vehicle 10. The position information being obtained as unsprung displacement z₁Indicates the moment of time when the unsprung displacement z was acquired₁Wheel of (2)11 (e.g., latitude and longitude). In fig. 2, the unsprung mass z registered in the forecast reference data 45 as the associated mass z₁c and one example of position information, illustrating the unsprung displacement "z₁cn "and position information" Xn, Yn "(n ═ 1, 2, 3 … …).

The ECU30 is connected to each of the front left wheel drive actuator 17FL, the front right wheel drive actuator 17FR, the rear left wheel drive actuator 17RL, and the rear right wheel drive actuator 17RR via a drive circuit (not shown).

The ECU30 predicts the unsprung position z based on the passage of each wheel 11 to be described later₁A target vibration damping control force Fct for damping the sprung mass 51 of each wheel 11 is calculated, and the active actuator 17 is controlled so that the vibration damping control force Fc generated by the active actuator 17 when each wheel 11 passes through the predicted position becomes the target vibration damping control force Fct.

Overview of basic predictive damping control

Hereinafter, an outline of the predictive vibration damping control common to the respective embodiments, which is executed by the running state control device 20, will be described. Fig. 3 shows a single wheel model of the vehicle 10.

The spring 52 corresponds to the suspension spring 16, the damper 53 corresponds to the shock absorber 15, and the actuator 54 corresponds to the active actuator 17.

In fig. 3, the mass of the sprung mass 51 is denoted as sprung mass m₂. The displacement in the vertical direction of the unsprung portion 50 described above is represented as unsprung portion displacement z₁. Further, the vertical displacement of the sprung portion 51 is represented as a sprung displacement z₂Sprung displacement z₂Is the vertical displacement of the spring 51 corresponding to the position of each wheel 11. The spring constant (equivalent spring constant) of the spring 52 is described as a spring constant K. The damping coefficient (equivalent damping coefficient) of the damper 53 is described as a damping coefficient C. The force generated by the actuator 54 is referred to as the control force Fc.

And, z₁And z₂Respectively, are described as dz₁And dz₂，z₁And z₂The second order time differential values of (A) are respectively described as ddz₁And ddz₂. In addition, z is₁And z₂The upward displacement is positive, and the upward force is positive with respect to the forces generated by the spring 52, the damper 53, the actuator 54, and the like.

In the single-wheel model of the vehicle 10 shown in fig. 3, the equation of motion relating to the vertical motion of the sprung portion 51 can be expressed by equation (1).

m₂ddz₂＝C(dz₁－dz₂)+K(z₁－z₂)－Fc……(1)

The damping coefficient C in equation (1) is assumed to be constant. However, since the actual damping coefficient changes according to the stroke speed of the suspension 13, the damping coefficient C may be set variably according to the time differential value of the stroke H, for example.

Also, in the case where the vibration on the spring 51 is completely eliminated by the vibration damping control force Fc (i.e., the sprung acceleration ddz₂Sprung velocity dz₂And sprung displacement z₂When each of the damping control forces Fc becomes zero), the damping control force Fc is expressed by equation (2).

Fc＝Cdz₁+Kz₁……(2)

Therefore, the damping control force Fc for reducing the vibration of the sprung mass 51 can be expressed by equation (3) with the control gain α. The control gain α is an arbitrary constant greater than 0 and equal to or less than 1.

Fc＝α(Cdz₁+Kz₁)……(3)

When equation (3) is applied to equation (1), equation (1) can be expressed by equation (4).

m₂ddz₂＝C(dz₁－dz₂)+K(z₁－z₂)－α(Cdz₁+Kz₁)……(4)

When the formula (4) is subjected to laplace transform and then arranged, the formula (4) is expressed by the formula (5). I.e. the unsprung displacement z₁Displacement z to the spring₂The transfer function of (2) is expressed by equation (5). In the formula (5), "s" is a laplace operator.

The value of the transfer function varies according to α according to equation (5), and when α is 1, the value of the transfer function is the smallest. Therefore, the target vibration damping control force Fct can be expressed by the following equation (6) corresponding to equation (3). The gain β in expression (6) is calculated₁Corresponding to the alpha Cs, gain beta₂Corresponding to α K.

Fct＝β₁×dz₁+β₂×z₁……(6)

Thus, the ECU30 acquires (pre-reads) the unsprung displacement z at the position (passing predicted position) that the wheel 11 will pass after in advance₁And by displacing the acquired unsprung mass by z₁The target damping control force Fct is calculated by applying equation (6). Then, the ECU30 determines the timing at which the wheel 11 passes through the passing predicted position (that is, the timing at which the unsprung mass z applied to equation (6) occurs)₁Timing of (c) to cause the actuator 54 to generate the damping control force Fc corresponding to the target damping control force Fct. If this is done, it is possible to reduce the unsprung mass z applied to equation (6) when the wheel 11 passes from the passing predicted position (i.e., when the unsprung mass z applied to equation (6) is generated)₁Time) of the vibration of the sprung 51.

The above is the vibration damping control of the sprung mass 51, and such vibration damping control of the sprung mass 51 is referred to as "predictive vibration damping control".

Note that, in the single wheel model described above, the mass of the unsprung portion 50 and the elastic deformation of the tire are ignored, and the road surface displacement z is assumed₀And unsprung displacement z₁The same is true. Thus, the road surface displacement z may also be used₀Instead of unsprung displacement z₁The same predictive vibration damping control is executed.

Predictive damping control of front and rear wheels

Next, predictive vibration damping control for the front wheels and the rear wheels common to the respective embodiments will be described with reference to fig. 4 to 6.

Fig. 4 shows the vehicle 10 at the present time tp traveling in the direction indicated by the arrow a1 at the vehicle speed V1. In the following description, the front wheels 11F and the rear wheels 11R are wheels on the same side, and the moving speeds of the front wheels 11F and the rear wheels 11R are assumed to be the same as the vehicle speed V1.

In fig. 4, a line Lt is a virtual time axis. Unsprung mass z on the course of movement of the front wheels 11F at the present, past and future times t₁A function z of a hypothetical time axis t, shown by the line Lt₁(t) represents. Thereby, the unsprung mass z of the position (ground point) pf0 of the front wheel 11F at the current time tp₁Is represented as z₁(tp). The unsprung mass z at the position pr0 at the current time tp of the rear wheel 11R₁Unsprung mass z of the front wheel 11F at time "tp-L/V1" earlier than the current time tp by "time (L/V1) taken for the front wheel 11F to move the wheelbase by L₁. Thereby, the unsprung displacement z of the rear wheel 11R at the current time tp₁Is represented as z₁(tp－L/V1)。

Predictive vibration damping control of front wheels 11F

The ECU30 determines the passing predicted position pf1 of the front wheel 11F that is later than the current time tp by the front wheel read-ahead time tpf (future). The front wheel read-ahead time tpf is set in advance to the time taken from when the ECU30 determines the passing predicted position pf1 until the front wheel active actuator 17F outputs the damping control force Fcf corresponding to the target damping control force Fcft.

The passing predicted position pf1 of the front wheel 11F is a position that is a front wheel read-ahead distance Lpf (V1 × tpf) from the position pf0 at the current time tp along the front wheel movement predicted route that is the route predicted that the front wheel 11F will move in the future. As will be described later in detail, the position pf0 is calculated based on the current position of the vehicle 10 acquired by the position information acquisition device 33.

The ECU30 acquires the unsprung position displacement through the predicted position pf1 as the unsprung position displacement z after determining that the front wheel passes through the predicted position pf1₁(tp + tpf). Then, the ECU30 calculates the unsprung mass z₁Time differential value dz of (tp + tpf)₁(tp + tpf). Note that, the acquisition of the time differential value of the unsprung displacement and the unsprung displacement of the front wheel by the predicted position differs depending on the embodiment,the gist of their acquisition will be explained later.

ECU30 responds by displacing unsprung mass by z₁(tp + tpf) and time differential value dz₁(tp + tpf) is applied to the following equation (7) corresponding to the above equation (6) to calculate the front wheel target vibration damping control force Fcft.

Fcft＝β₁f×dz₁+β₂f×z₁……(7)

Furthermore, the ECU30 sends a control command including the target damping control force Fcft to the front wheel active actuator 17F so that the front wheel active actuator 17F generates a damping control force Fcf corresponding to the target damping control force Fcft.

As shown in fig. 5, the front-wheel active actuator 17F generates the damping control force Fcf corresponding to the target damping control force Fcft at "time tp + tpf" that is later than the current time tp by the front-wheel read-ahead time tpf (i.e., at the timing at which the front wheel 11F actually passes from the passing predicted position pf 1). Thus, the front wheel active actuator 17F can generate the unsprung displacement z absorbed by the predicted passing position pf1 of the front wheel 11F at an appropriate timing₁And the generated damping control force Fcf of the exciting force, thereby appropriately reducing the vibration of the sprung mass 51.

Predictive vibration damping control of rear wheel 11R

The ECU30 determines the passing predicted position pr1 of the rear wheels 11R later than the current time tp by the rear wheel read-ahead time tpr (future). The rear wheel read-ahead time tpr is set in advance as the time taken from when the ECU30 determines the passing predicted position pr1 until the rear wheel active actuator 17R outputs the damping control force Fcr corresponding to the target damping control force Fcrt.

When the front wheel drive actuator 17F and the rear wheel drive actuator 17R are different drive actuators, the front wheel read-ahead time tpf and the rear wheel read-ahead time tpr are preset to different values. In the case where the front wheel active actuator 17F and the rear wheel active actuator 17R are the same active actuator, the front wheel read-ahead time tpf and the rear wheel read-ahead time tpr are set in advance to the same value.

The ECU30 predicts the movement of the rear wheels 11R in a case where it is assumed that the rear wheels 11R advance along the same route as the front wheels 11FThe position of the measured line away from the position of the current time tp by the rear wheel read-ahead distance Lpr (V1 × tpr) is determined as the passing predicted position pr 1. The unsprung mass z passing through the predicted position pr1₁Unsprung mass z at the rear wheel read-ahead time tpr later than the "time (tp-L/V1) at which the front wheel 11F is located at the position pr0 of the current time of the rear wheel 11R" (tp-L/V1 + tpr)₁。

Thus, the ECU30 acquires the unsprung position displacement of the rear wheel passing through the predicted position pr1 as the unsprung position displacement z₁(tp-L/V1 + tpr). Then, the ECU30 calculates the unsprung mass z₁Time differential value dz of (tp-L/V1 + tpr)₁(tp-L/V1 + tpr). Note that, since acquisition of the unsprung displacement at the predicted position of the rear wheel and the time differential value of the unsprung displacement differs depending on the embodiment, the gist of acquisition thereof will be described later.

ECU30 responds by displacing unsprung mass by z₁(tp-L/V1 + tpr) and time derivative value dz₁(tp-L/V1 + tpr) is applied to the following equation (8) corresponding to the above equation (6) to calculate the rear wheel target vibration damping control force Fcrt.

Fcrt＝β₁r×dz₁+β₂r×z₁……(8)

Then, the ECU30 transmits a control command including the target damping control force Fcrt to the rear wheel active actuator 17R so that the rear wheel active actuator 17R generates the damping control force Fcr corresponding to the target damping control force Fcrt.

As shown in fig. 6, the rear wheel active actuator 17R generates the damping control force Fcr corresponding to the target damping control force Fcrt at "time tp + tpr" that is later than the current time tp by the rear wheel read-ahead time tpr. Thus, the rear wheel active actuator 17R can generate the unsprung displacement z absorbed by the predicted passing position pr1 of the rear wheel 11R at an appropriate timing₁And the vibration damping control force Fcr of the exciting force is generated, thereby appropriately reducing the vibration of the sprung mass 51.

Roll control

Next, roll control on a spring common to the respective embodiments will be described. During turning of the vehicle 10, the ECU30 calculates an estimated lateral acceleration Gyh of the vehicle in accordance with a technique known in the art based on the vehicle speed V1 and the steering angle θ. The estimated lateral acceleration Gyh has a positive value when the vehicle turns left. The ECU30 calculates the target roll control force Frit based on the absolute value of the estimated lateral acceleration Gyh so that the greater the absolute value of the estimated lateral acceleration Gyh of the vehicle, the greater the magnitude of the anti-roll moment due to the control force F generated by the active actuator 17 of each wheel.

The ECU30 controls the active actuators such that the roll control forces Fri generated by the active actuators 17 of the respective wheels become the corresponding target roll control forces Fri, respectively. Note that i is fl, fr, rl, and rr, and refers to the left front wheel, the right front wheel, the left rear wheel, and the right rear wheel, respectively.

Running state control routine of the first embodiment

The CPU of the ECU30 executes a vibration damping control routine shown in the flowcharts of fig. 7 and 8 each time a prescribed time elapses. Unless otherwise specified, "CPU" refers to the CPU of the ECU 30.

When the predetermined timing is reached, the CPU starts processing from step 700 in fig. 7 to execute step 710 to step 780, and then proceeds to step 790 to once end the present routine.

Step 710: the GNSS signals received by the position information acquiring device 33 contain information relating to the moving speed, and the CPU determines the vehicle speed V1 based on the GNSS signals. Then, the CPU calculates an estimated lateral acceleration Gyh of the vehicle in accordance with a technique known in the art based on the vehicle speed V1 and the steering angle θ.

Step 720: the CPU calculates the target roll control force Frit by referring to the map shown in fig. 9 based on the absolute value of the estimated lateral acceleration Gyh. In fig. 9, the solid line is a map when estimated lateral acceleration Gyh is positive, and the broken line is a map when estimated lateral acceleration Gyh is negative. The ratio of the target roll control forces of the front and rear wheels preferably corresponds to a ratio of a distance Lr between the center of gravity on the spring and the axle of the rear wheel and a distance Lf between the center of gravity on the spring and the axle of the front wheel.

As shown in fig. 9, the target roll control force Frit is 0 when the absolute value of the estimated lateral acceleration Gyh is Gyh0 (a constant equal to or greater than 0) or less. When the absolute value of the estimated lateral acceleration Gyh is greater than Gyh0, the magnitude of the target roll control force Frit increases as the absolute value of the estimated lateral acceleration Gyh increases. Thus, the target roll control force Frit is calculated as: the larger the absolute value of the estimated lateral acceleration Gyh of the vehicle is, the larger the magnitude of the anti-roll moment caused by the roll control force generated by the active actuator 17 of each wheel is.

Step 730: the CPU determines whether or not the switch 35 is on, and proceeds the control to step 750 when the switch 35 is on, and proceeds the control to step 740 when the switch 35 is off.

Step 740: the CPU sets the target damping control forces Fcit for all the wheels to 0. Step 750: the CPU calculates the target damping control force Fcit for each wheel in accordance with the calculation control routine shown in fig. 8.

Step 760: the CPU calculates a correction coefficient Ac for the target vibration damping control force Fcit by referring to the map shown in fig. 10 based on the estimated lateral acceleration Gyh of the vehicle. As shown in fig. 10, the correction coefficient Ac is 1 when the absolute value of the estimated lateral acceleration Gyh is Gyh0 or less. The correction coefficient Ac is calculated when the absolute value of the estimated lateral acceleration Gyh is greater than Gyh0 as: the correction coefficient Ac becomes a smaller positive value as the absolute value of the estimated lateral acceleration Gyh becomes larger.

Step 770: the CPU calculates the target control force Fit of the active actuator 17 for each wheel according to the following equation (9).

Fit＝Frit+Ac×Fcit……(9)

Step 780: the CPU transmits a control command including the target control force Fit to the active actuator 17 of each wheel, thereby performing control so that the control force F generated by each active actuator becomes the target control force Fit. Each active actuator outputs a damping control force corresponding to the product of the correction coefficient Ac and the target damping control force Fcit at the timing at which each wheel 11 passes through the corresponding predicted passing position.

Calculation of target vibration damping control force Fcit in step 750

Step 810: the CPU acquires information relating to the current position of the vehicle 10 from the position information acquisition device 33, and determines (acquires) the current position of each wheel 11, the vehicle speed V1, and the traveling direction Td of the vehicle 10.

More specifically, the CPU maps the previous current position and the present current position to road map information included in the map database, and determines a direction from the previous current position toward the present current position as the traveling direction Td of the vehicle 10. The last current position refers to the current position of the vehicle 10 acquired by the CPU in step 710 of the present routine executed last time. The present current position refers to the current position of the vehicle 10 acquired by the CPU in the present step 710.

The ROM of the ECU30 stores positional relationship data indicating the relationship between the mounting position of the GNSS receiver on the vehicle 10 and the position of each wheel 11 in advance. Since the current position of the vehicle 10 acquired from the position information acquiring device 33 corresponds to the mounting position of the GNSS receiver, the CPU determines the current position of each wheel 11 by referring to the current position of the vehicle 10, the traveling direction Td of the vehicle 10, and the above-described positional relationship data. Also, the CPU determines the vehicle speed V1 based on the GNSS signal.

Step 820: the CPU determines the front-wheel movement predicted route and the rear-wheel movement predicted route as described below.

The CPU determines the routes predicted to be moved by the left and right front wheels 11FL and 11FR when the vehicle 10 travels in the traveling direction Td as the left and right front wheel movement predicted routes, based on the current position of each wheel 11, the traveling direction Td of the vehicle 10, and the above-described positional relationship data.

The left and right rear wheel movement predicted routes include "a first predicted route from the current positions of the left and right rear wheels 11RL and 11RR to the current positions of the front wheels 11FL and 11FR, respectively" and "a second predicted route on the traveling direction side of the vehicle 10 with respect to the current positions of the front wheels 11FL and 11 FR". Therefore, the CPU determines the paths on which the left and right front wheels 11FL and 11FR actually move from the current positions of the rear wheels 11RL and 11RR to the current positions of the left and right front wheels as the first predicted route. Then, the CPU determines the left and right predicted front wheel movement routes as second predicted routes for the left and right rear wheels 11RL and 11RR, respectively.

As described above, the CPU calculates the front wheel read-ahead distance Lpf by multiplying the vehicle speed V1 by the front wheel read-ahead time tpf, and calculates the rear wheel read-ahead distance Lpr by multiplying the vehicle speed V1 by the rear wheel read-ahead time tpr. Further, the CPU determines the front wheel passing predicted position pf1 and the rear wheel passing predicted position pr 1.

More specifically, the CPU determines the positions of the left and right front wheels 11FL and 11FR after having advanced the front wheel pre-read distance Lpf along the left and right front wheel passing predicted route from their current positions as the left and right front wheel passing predicted positions pfl1, pfr 1. Further, the CPU determines the positions of the left and right rear wheels 11RL and 11RR which have advanced the rear wheel read-ahead distance Lpr from their current positions along the left and right rear wheel passing predicted routes as the left and right rear wheel passing predicted positions prl1, prr 1.

Step 830: the CPU acquires unsprung displacement z of the front wheel passing predicted position from unsprung displacement in the "preparation section" acquired from forecast reference data 45 of cloud 40 in advance₁ci information.

The preparation section is a section having the front wheel passing predicted position pf1 as a starting point when reaching the end point of the preparation section, and having a position separated from the front wheel passing predicted position pf1 by a predetermined preparation distance along the front wheel movement predicted route as an end point. The preparatory distance is set to a sufficiently larger value than the front wheel read-ahead distance Lpf.

Step 840: CPU by unsprung displacement z₁ci time-differentiating to calculate the unsprung velocity dz₁ci。

Step 850: CPU based on unsprung speed dz₁ci and unsprung displacement z₁ci calculates target damping control forces Fcflt, Fcfrt, Fcrlt, Fcrrt of the active actuators for the front left and right wheels by the following equations (10) and (11) corresponding to the above equations (7) and (8), respectively.

Fcit＝β₁f×dz₁ci+β₂f×z₁ci……(10)

Fcit＝β₁r×dz₁ci+β₂r×z₁ci……(11)

The gains β in expressions (10) and (11) are₁f、β₂f and a gain beta₁r、β₂r are respectively expressed as values different from each other. This is because it is considered that the damping coefficients Cf and Cr of the shock absorbers for the front and rear wheels are different from each other, and the spring constants Kf and Kr of the suspensions for the front and rear wheels are different from each other.

As can be understood from the above, according to the first embodiment, the ECU30 of the running state control device 20 performs the reduction correction of the target damping control force anticipating the damping control when the anticipating damping control and the rolling control are simultaneously executed. Thus, even when the predictive vibration damping control and the roll control are executed simultaneously, the possibility of the roll on the spring being deteriorated by the vibration damping control force of the predictive vibration damping control can be reduced as compared with the case where the reduction correction of the target vibration damping control force of the predictive vibration damping control is not performed.

In particular, according to the first embodiment, the correction coefficient Ac for reduction-correcting the target damping control force for the predictive damping control is calculated as: the correction coefficient Ac is smaller as the absolute value of the estimated lateral acceleration Gyh as the roll index value is larger. Therefore, the higher the possibility that the roll on the spring becomes larger, the larger the amount by which the control force of the vibration damping control is reduced, and therefore, the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control can be appropriately reduced as compared with the case where the correction coefficient Ac is constant and the reduction correction amount of the target vibration damping control force is constant.

Further, according to the first embodiment, the roll index value is the estimated lateral acceleration Gyh that changes earlier in time than the actual lateral acceleration of the vehicle. Thus, the delay of the decrease correction of the target vibration damping control force with respect to the change in the centrifugal force of the vehicle that changes the roll amount on the spring can be reduced as compared with the case where the roll index value is the actual lateral acceleration of the vehicle.

In the first embodiment, the correction coefficient Ac for reducing and correcting the target vibration damping control force for predictive vibration damping control is calculated based on the estimated lateral acceleration Gyh as the roll index value, and the target vibration damping control force is reduced and corrected by multiplying the target vibration damping control force by the correction coefficient Ac. However, the reduction correction amount Δ Fcit of the target damping control force Fcit may be calculated based on the roll index value and the target damping control force, and the target control force Fit may be calculated by subtracting the reduction correction amount Δ Fcit from the sum of the target roll control force Frit and the target damping control force Fcit.

In the first embodiment, the estimated lateral acceleration Gyh as the roll index value is calculated based on the vehicle speed V1 and the steering angle θ. However, the estimated lateral acceleration Gyh may be calculated as the product of the yaw rate of the vehicle detected by a yaw rate sensor or the yaw rate of the vehicle calculated based on the wheel speeds of the left and right wheels and the vehicle speed V1.

Second embodiment

In the vehicle running state control device 20 of the second embodiment shown in fig. 11, a lateral acceleration sensor 36 and a anticipation sensor 37 are provided in place of the steering angle sensor 31 and the vehicle speed sensor 32 in the vehicle-mounted device 21. The lateral acceleration sensor 36 and the anticipation sensor 37 are connected to the ECU 30. The lateral acceleration sensor 36 detects an actual lateral acceleration Gy of the vehicle 10, which becomes a positive value when the vehicle turns left.

The anticipatory sensor 37 may be any anticipatory sensor known in the art as long as it can acquire a value indicating a vertical displacement of a road surface in front of the vehicle 10 (referred to as "road surface displacement"), such as a camera sensor, a LIDAR, and a radar. The ECU30 functions as road surface displacement related information acquisition means for acquiring road surface displacement related information in front of each wheel based on the detection result of the anticipatory sensor in cooperation with the anticipatory sensor 37, which is a vehicle-mounted sensor. Thus, in this embodiment, the device 22 outside the vehicle, the position information acquiring device 33, and the wireless communication device 34 may be omitted.

As shown in fig. 11, the anticipatory sensor 37 is attached to, for example, the inner surface of the upper end portion of the center in the vehicle width direction of the windshield 10b of the vehicle 10, and detects the target position Po ahead of the anticipatory distance Lpre from the front wheel 11F and the target position Po ahead thereofPeripheral road surface displacement z₀. It is noted that the predicted distance Lpre is preferably larger than a front wheel read-ahead distance Lpf (described later) when the vehicle speed of the vehicle 10 is a rated maximum vehicle speed. In fig. 11, only one anticipation sensor 37 is shown, but a pair of anticipation sensors corresponding to the left and right front wheels may be provided.

Running state control routine of second embodiment

The running state control is executed by the ECU30 at predetermined intervals in accordance with a running state control routine shown in the flowchart of fig. 12, as in the first embodiment.

When the predetermined timing is reached, the CPU of the ECU30 starts processing from step 1200 of fig. 12 to execute step 1210 to step 1280, and then proceeds to step 1290 to once end the routine.

Step 1210: the CPU reads in the actual lateral acceleration Gy of the vehicle 10 detected by the lateral acceleration sensor 36.

Step 1220: the CPU calculates the target roll control force Frit by referring to the map shown in fig. 14 based on the absolute value of the actual lateral acceleration Gy. In fig. 14, the solid line is a map when the actual lateral acceleration Gy is positive, and the broken line is a map when the actual lateral acceleration Gy is negative. The ratio of the target roll control forces of the front and rear wheels preferably corresponds to a ratio of a distance Lr between the center of gravity on the spring and the axle of the rear wheel and a distance Lf between the center of gravity on the spring and the axle of the front wheel.

As shown in fig. 14, the target roll control force Frit is 0 when the absolute value of the actual lateral acceleration Gy is Gy0 (a constant equal to or greater than 0) or less. In the case where the absolute value of the actual lateral acceleration Gy is larger than Gy0, the magnitude of the target rolling control force Frit becomes larger as the absolute value of the actual lateral acceleration Gy is larger. Thus, the target roll control force Frit is calculated as: the larger the absolute value of the actual lateral acceleration Gy of the vehicle is, the larger the magnitude of the anti-roll moment caused by the roll control force generated by the active actuator 17 of each wheel is.

Step 1230: the CPU determines whether or not the switch 35 is on, and proceeds to step 1250 when the switch 35 is on, and proceeds to step 1240 when the switch 35 is off.

Step 1240: the CPU sets the target damping control forces Fcit for all the wheels to 0 and sets the correction coefficient Ar to 1.

Step 1250: the CPU calculates the target damping control force Fcit for each wheel in accordance with the calculation control routine shown in fig. 13.

Step 1260: the CPU calculates a correction coefficient Ar for the target roll control force Frit by referring to the map shown in fig. 15 based on the actual lateral acceleration Gy of the vehicle. As shown in fig. 15, the correction coefficient Ar is 1 when the absolute value of the actual lateral acceleration Gy is Gy0 or less. The correction coefficient Ar is calculated as: the correction coefficient Ar becomes a positive value that increases as the absolute value of the actual lateral acceleration Gy increases.

Step 1270: the CPU calculates the target control force Fit of the active actuator 17 for each wheel according to the following equation (12).

Fit＝Ar×Frit+Fcit……(12)

Calculation of target damping control force Fcit in step 1250

Step 1310: the CPU acquires information on the current position of the vehicle 10 from the position information acquiring device 33 in the same manner as step 810, and determines (acquires) the current position of each wheel 11, the vehicle speed V1, and the traveling direction Td of the vehicle 10.

Step 1320: the CPU determines the front-wheel movement predicted route and the rear-wheel movement predicted route in the same manner as step 820.

In step 1330, the CPU acquires unsprung mass z of each wheel passing through the predicted position based on the road surface displacement in front of the vehicle detected by the anticipation sensor 37₁si. In this case, the road surface displacement z of the wheel passing the predicted position detected by the anticipatory sensor 37 may be acquired₀si as unsprung displacement z₁si. Further, the road surface displacement in front of the vehicle detected by the anticipation sensor 37 may be temporarily stored in the RAM, and the road surface displacement z of the front wheel passing predicted position may be determined based on the stored road surface displacement₀si，Obtaining the road surface displacement z₀si as unsprung displacement z₁si。

Step 1340: CPU by unsprung displacement z₁si is time differentiated to calculate the unsprung velocity dz₁si。

Step 1350: CPU based on unsprung speed dz₁si and unsprung mass z₁si, the target vibration damping control forces Fcit of the active actuators 17 for the right and left front wheels and the right and left rear wheels are calculated by the following equations (13) and (14) corresponding to the above equations (7) and (8), respectively.

Fcit＝β₁f×dz₁si+β₂f×z₁si……(13)

Fcit＝β₁r×dz₁si+β₂r×z₁si……(14)

As can be understood from the above, according to the second embodiment, the ECU30 of the running state control device 20 increases the roll control force by correcting the increase in the target roll control force for the roll control when executing the predictive vibration damping control and the roll control simultaneously. Thus, even when the predictive vibration damping control and the roll control are executed simultaneously, the possibility of the roll on the spring being deteriorated by the control force of the predictive vibration damping control can be reduced as compared with the case where the target roll control force of the roll control is not corrected to be increased.

In particular, according to the second embodiment, the correction coefficient Ar for increasing-correcting the target roll control force for roll control is calculated as: the correction coefficient Ar becomes larger as the absolute value of the actual lateral acceleration Gy as the roll index value becomes larger. Therefore, the higher the possibility that the roll on the spring becomes larger, the larger the amount by which the effect of the roll control increases, and therefore, the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control can be appropriately reduced as compared with the case where the correction coefficient Ar is constant and the increase correction amount of the target roll control force is constant.

Further, according to the second embodiment (and the third embodiment described later), the roll index value is not an estimated value of the lateral acceleration of the vehicle, but an actual lateral acceleration Gy detected by the lateral acceleration sensor. Thus, the roll control force generated by the estimation and the error in the increase control thereof can be reduced as compared with the case where the roll index value is the estimated lateral acceleration of the vehicle.

In the second embodiment, the correction coefficient Ar for increasing and correcting the target roll control force for roll control is calculated based on the actual lateral acceleration Gy as the roll index value, and the target roll control force is increased and corrected by multiplying the target roll control force by the correction coefficient Ar. However, the increase correction amount Δ Frit of the target roll control force Frit may be calculated based on the roll index value and the target roll control force, and the target control force Fit may be calculated as the sum of the target roll control force Frit, the target vibration damping control force Fcit, and the increase correction amount Δ Frit.

Third embodiment

Running state control routine of third embodiment

The running state control in the third embodiment is executed by the ECU30 at predetermined intervals in accordance with a running state control routine shown in the flowchart of fig. 16, as in the second embodiment.

When the predetermined timing is reached, the CPU of the ECU30 starts the process from step 1600 in fig. 16 to execute step 1610 to step 1680, and then proceeds to step 1690 to once end the routine.

As can be seen from comparison between fig. 16 and fig. 12, steps 1610 to 1650 and steps 1670 and 1680 are performed in the same manner as steps 1210 to 1250 and steps 1270 and 1280, respectively, in the second embodiment.

In step 1660, the CPU calculates an index value Fca of the vibration damping control indicating the magnitude of the control force of the vibration damping control, and calculates a correction coefficient Ar for the target roll control force Frit by referring to the map shown in fig. 17 based on the index value Fca. As shown in fig. 17, the correction coefficient Ar is 1 when the index value Fca is equal to or less than Fca0 (constant equal to or greater than 0). When the index value Fca is larger than Fca0, the correction coefficient Ar is calculated as: the correction coefficient Ar becomes a positive value that increases as the index value Fca increases.

The index value Fca of the vibration damping control may be a maximum value of a moving average value Fcita of absolute values of the target vibration damping control forces Fcit of the four wheels within a predetermined time t0 (normal number) set in advance, or a maximum value of a peak-to-peak value PFcit of the target vibration damping control forces Fcit of the four wheels within a predetermined time t 0. The index value Fca of the vibration damping control may be a maximum value of a moving average of absolute values of the unsprung displacements of the four wheels within a predetermined time t0, or a maximum value of peak-to-peak values of the unsprung displacements of the four wheels within a predetermined time t 0.

As can be understood from the above, according to the third embodiment, when the predictive vibration damping control and the roll control are simultaneously executed, the ECU30 of the running state control device 20 increases the target roll control force of the roll control by increasing the correction to increase the roll control force, as in the second embodiment. Thus, even when the predictive vibration damping control and the roll control are executed simultaneously, the possibility of the roll on the spring being deteriorated by the control force of the predictive vibration damping control can be reduced as compared with the case where the target roll control force of the roll control is not corrected to be increased.

In particular, according to the third embodiment, the correction coefficient Ar for increase-correcting the target roll control force for roll control is calculated as: the correction coefficient Ar is larger as the index value Fca of the vibration damping control indicating the magnitude of the control force of the vibration damping control is larger. Thus, the greater the magnitude of the control force of the vibration damping control, the higher the possibility that the roll on the spring is deteriorated by the control force of the vibration damping control, the greater the amount by which the effect of the roll control is increased. Therefore, the possibility of the roll on the spring being deteriorated by the control force of the vibration damping control can be appropriately reduced, as compared with the case where the correction coefficient Ar is constant and the increase correction amount of the target roll control force is constant.

In the third embodiment, the correction coefficient Ar for increasing and correcting the target roll control force for roll control is calculated based on the index value Fca of vibration damping control, and the target roll control force is increased and corrected by multiplying the target roll control force by the correction coefficient Ar. However, the increase correction amount Δ Frit of the target roll control force Frit may be calculated based on the index value Fca of the vibration damping control and the target roll control force, and the target control force Fit may be calculated as the sum of the target roll control force Frit, the target vibration damping control force Fcit, and the increase correction amount Δ Frit.

[ modified examples ]

The roll index value in the first embodiment described above may be replaced with the actual lateral acceleration Gy detected by the lateral acceleration sensor, and conversely, the roll index values in the second and third embodiments may be replaced with the estimated lateral acceleration Gyh of the vehicle. In the first to third embodiments, the roll index value may be replaced with, for example, a roll angle on a spring calculated based on a stroke detected by a stroke sensor incorporated in a suspension of each wheel.

In the first embodiment described above, the unsprung mass z of the wheel passing through the predicted position₁ci and unsprung velocity dz₁ci based on unsprung displacement z obtained from cloud 40₁To obtain the final product. However, the unsprung mass z of the wheel passing the predicted position in the first embodiment₁ci and unsprung velocity dz₁ci may be determined based on the road surface displacement in front of the vehicle detected by the anticipation sensor 37, as in the second and third embodiments.

Conversely, in the second and third embodiments described above, the unsprung mass z of the wheel passing through the predicted position₁si and unsprung velocity dz₁si is obtained based on the road surface displacement in front of the vehicle detected by the anticipation sensor 37. However, the unsprung mass z of the wheel passing the predicted position in the second and third embodiments₁si and unsprung velocity dz₁si may be based on unsprung mass z acquired from cloud 40 in the same manner as in the first embodiment₁To obtain the final product.

The unsprung displacement and the unsprung velocity at the wheel passing predicted positions in the first to third embodiments may be calculated by a technique known in the art based on the sprung vertical acceleration and the suspension stroke, or the unsprung vertical acceleration at the position of each wheel. Further, the unsprung displacement and the unsprung velocity of the wheel by the predicted position may be calculated using an observer known in the art based on at least one of the sprung vertical acceleration, the suspension stroke, and the unsprung vertical acceleration at the position of each wheel.

In the first embodiment described above, the target damping control force for the predictive damping control is corrected to be decreased when the predictive damping control and the roll control are executed simultaneously. In the second and third embodiments, the target roll control force of the roll control is corrected to be increased when the predictive vibration damping control and the roll control are simultaneously executed. However, in the first to third embodiments, when the prospective damping control and the rolling control are executed simultaneously, the target damping control force for the prospective damping control may be corrected to be decreased and the target rolling control force for the rolling control may be corrected to be increased.

In the first to third embodiments, the target roll control force Frit of the roll control on the spring is calculated based on the estimated lateral acceleration Gyh or the actual lateral acceleration Gy of the vehicle as the roll index value. The roll on the spring is feedforward controlled based on the estimated lateral acceleration Gyh or the actual lateral acceleration Gy. However, the roll on the spring may also be feedback controlled based on the deviation of the roll angle on the spring from the target roll angle. In this case, the increase correction of the target roll control force of the roll control may be achieved by the increase correction of the feedback control amount.

Also, the roll on the spring can be controlled by both feedforward control and feedback control. In this case, the increase correction of the target roll control force for the roll control can be achieved by the increase correction of one or both of the feedforward control amount and the feedback control amount.

While the present invention has been described in detail with reference to the specific embodiments, it is apparent to those skilled in the art that the present invention is not limited to the embodiments described above, and various other embodiments can be implemented within the scope of the present invention.

For example, in the first to third embodiments described above, the switch 35 is provided, and the predictive vibration damping control is executed when the switch 35 is on. However, the switch 35 may be omitted, the steps 730 and 740 in the first embodiment, the steps 1230 and 1240 in the second embodiment, and the steps 1630 and 1640 in the third embodiment may be omitted.

In the first embodiment described above, it is anticipated that the reference data 45 need not be stored in the storage device 44 of the cloud 40, but may be stored in the storage device 30 a.

When the travel route of the vehicle 10 is determined in advance, the CPU may download the forecast reference data 45 of the travel route from the cloud 40 in advance before the vehicle 10 starts traveling the travel route, and store the data in the storage device 30 a.

In the forecast reference data 45, the unsprung speed dz may be stored in association with the position information and the vehicle speed information₁ci instead of unsprung displacement z₁. In this case, for example, in step 750 shown in fig. 7, the CPU acquires the unsprung speed dz₁ci by comparing the acquired unsprung speed dz₁ci integral to calculate the unsprung mass z₁ci。

The processing of calculating the target damping control force Fcrt for the rear wheels 11R in the first to third embodiments is not limited to the above example. For example, the CPU may also make the unsprung displacement z based on the current position of the front wheel 11F at the current time tp₁The target damping control force Fcrt is calculated in advance, and a control command including the target damping control force Fcrt is transmitted to the rear wheel active actuator 17R at a timing delayed by a time (L/V-tpr) from the current time tp. That is, the CPU may transmit a control command including the target vibration damping control force Fcrt to the rear wheel active actuator 17R at a timing when the rear wheel 11R reaches a position closer to the front-rear wheel read-ahead distance Lpr than the current position of the front wheel 11F.

Further, the CPU determines a rear wheel movement prediction route based on the current position of the rear wheels 11R, the traveling direction Td of the vehicle 10, and the positional relationship data independently of the front wheel movement prediction route, and determines a position a rear wheel read-ahead distance Lpr from the rear wheel along the rear wheel movement prediction route as a rear wheel passing prediction position. Then, the CPU obtains the unsprung displacement z of the rear wheel through the predicted position₁Based on the acquired unsprung displacement z₁The target damping control force Fcrt for the rear wheels 11R is calculated.

The vehicle speed V1 and the traveling direction Td are acquired based on the current position of the vehicle 10 acquired by the GNSS receiver, but are not limited thereto. For example, the running state control device 20 may include a "wheel speed sensor and a steering angle sensor" not shown, the wheel speed sensor detecting the rotation speed of the wheel 11, and the CPU calculating the vehicle speed V1 based on the rotation speed of the wheel 11. Alternatively, a yaw rate sensor for detecting a yaw rate of the vehicle 10 may be provided, and the CPU may acquire the traveling direction Td based on the yaw rate and the vehicle speed V1.

The suspensions 13FL to 13RR may be any type of suspension as long as they respectively allow the wheels 11FL to 11RR and the vehicle body 10a to be displaced in the up-down direction relative to each other. Further, the suspension springs 16FL to 16RR may be any of compression coil springs, air springs, and the like.

In each of the above embodiments, the active actuators 17FR to 17RR are provided corresponding to the respective wheels 11, but one active actuator 17 may be provided for at least one wheel 11. For example, the vehicle 10 may be provided with only one of the front wheel drive actuator 17F and the rear wheel drive actuator 17R.

In the above-described embodiment and the above-described modification, the active actuator 17 is used as the control force generation device, but the present invention is not limited thereto. That is, the control force generating device may be an actuator that can generate the control force in the vertical direction for damping the sprung portion 51 so as to be adjustable based on a control command including the target control force.

The control force generating device may be an active stabilizer (not shown). The active stabilizer device comprises a front wheel active stabilizer and a rear wheel active stabilizer. In the front wheel active stabilizer, when a control force in the vertical direction (left front wheel control force) is generated between the sprung portion 51 and the unsprung portion 50 corresponding to the left front wheel 11FL, a control force in the opposite direction to the left front wheel control force (right front wheel control force) is generated between the sprung portion 51 and the unsprung portion 50 corresponding to the right front wheel 11 FR. Similarly, in the rear wheel active stabilizer, when a control force in the vertical direction (left rear wheel control force) is generated between the sprung portion 51 and the unsprung portion 50 corresponding to the left rear wheel 11RL, a control force in the opposite direction to the left rear wheel control force (right rear wheel control force) is generated between the sprung portion 51 and the unsprung portion 50 corresponding to the right rear wheel 11 RR. The structure of the above-described active stabilizer device is well known and is incorporated in the present specification by referring to japanese patent application laid-open No. 2009-96366. The active stabilizer device may include at least one of a front wheel active stabilizer and a rear wheel active stabilizer.

The control force generating device may be a device that generates the control force F in the vertical direction by the geometry of the suspensions 13FL to 13RR by increasing or decreasing the braking/driving force for each wheel 11 of the vehicle 10. The structure of such a device is well known, and is incorporated in the present specification by referring to japanese patent application laid-open No. 2016-107778 and the like. The ECU30 calculates a braking/driving force that generates the control force F corresponding to the target control force Ft by a well-known method.

Such devices include a driving device (for example, an in-wheel motor) that applies driving force to each wheel 11 and a brake device (brake device) that applies braking force to each wheel 11. The drive device may be a motor, an engine, or the like that applies a driving force to one or four wheels of the front wheels and the rear wheels. The control force generating device may include at least one of a driving device and a braking device.

Further, the control-force generating device may be the damping-force variable type shock absorbers 15FL to 15 RR. In this case, the ECU30 controls the damping coefficients C of the shock absorbers 15FL to 15RR in such a manner that the damping forces of the shock absorbers 15FL to 15RR change by values corresponding to the target damping control force Fct.