JP5724524B2 - Car body vibration control device and car body vibration control method - Google Patents

Description

  The present invention relates to a vehicle body vibration control device and a vehicle body vibration control method.

Conventionally, as this type of technology, for example, there is a technology described in Patent Document 1.
In the technique described in Patent Document 1, the driver's steering state is estimated based on the steering speed, lateral G, yaw angular acceleration, and the like, and vibrations of the vehicle body (for example, bounce of the vehicle body, The feedback gain used for damping control that suppresses (pitch behavior) is increased. As a result, the absolute value of the feedback value is increased, the bounce and pitch behavior of the vehicle body are more reliably suppressed, and the steering stability of the vehicle is improved.

JP 2008-179277 A

However, although the technique described in Patent Document 1 can suppress the vibration of the vehicle body, that is, the fluctuation of the wheel load, the roll behavior increases due to an increase in the absolute value of the feedback value, which makes the driver feel uncomfortable. There was a possibility.
This invention pays attention to the above points, and makes it a subject to make it possible to suppress that a roll behavior increases, when steering operation is performed.

In order to solve the above-described problems, the present invention controls the drive torque in a direction that suppresses fluctuations in required drive torque and components caused by road surface disturbance among components constituting the sprung behavior of the vehicle body. Further, the drive torque is controlled in a direction that promotes the fluctuation of the front wheel load, which is a component caused by the turning resistance among the components constituting the sprung behavior of the vehicle body. Based on the required drive torque, road disturbance, and turning resistance, the rotational motion of the vehicle body about the pitch axis and the vertical motion in the bounce direction are estimated as the sprung behavior of the vehicle body. Based on the estimated rotational movement of the vehicle body around the pitch axis and the vertical movement in the bounce direction, the driving torque applied by the torque adding means is controlled in a direction that promotes the fluctuation of the physical quantity related to the front wheel load, which is a component caused by turning resistance. By doing so, roll suppression control is performed to suppress the roll behavior of the vehicle body. Further, as the roll suppression control, after the driver starts the steering operation, the driving torque is reduced before the roll angular velocity generated by the steering operation reaches the peak.

According to such a configuration, before starting the steering operation, the wheel load fluctuation is suppressed by controlling the driving torque in a direction to suppress the fluctuation of the component caused by the driver's requested driving torque and road surface disturbance. can do. Also, when the steering operation is started, by controlling the driving torque in the direction that promotes the fluctuation of the front wheel load caused by the turning resistance applied to the wheel, the nose down behavior can be promoted, the wheel load of the front wheel can be increased, the steering Responsiveness can be improved. Further, by improving the steering response while suppressing the fluctuation of the wheel load, the change in the lateral G can be moderated, and the roll behavior can be suppressed when the steering operation is performed. In addition, a physical quantity that can be controlled by driving torque can be estimated as the sprung behavior of the vehicle body. Therefore, the roll behavior can be appropriately suppressed by controlling the driving torque. Furthermore, by reducing the drive torque before the roll angular velocity reaches its peak, the nose-down behavior can be promoted, the wheel load of the front wheels can be increased, and the steering response can be improved.

It is a conceptual diagram showing the schematic structure of the vehicle of 1st Embodiment. It is a block diagram showing the schematic structure of the vehicle of a 1st embodiment. It is a block diagram showing the structure of the program which a microprocessor performs. 3 is a block diagram illustrating a configuration of a driving power wheel system vibration control unit 16. FIG. 3 is a flowchart showing the operation of a driving power wheel system vibration control unit 16. 3 is a block diagram illustrating a configuration of a suspension stroke calculation unit 21. FIG. 3 is a flowchart showing the operation of a suspension stroke calculation unit 21. It is a figure for demonstrating the calculation method of the stroke amount of a suspension. It is a graph showing a front-wheel suspension geometry characteristic. It is a graph showing a rear-wheel suspension geometry characteristic. 4 is a diagram for explaining a vehicle model 26. FIG. FIG. 6 is an explanatory diagram for explaining the operation of a torque command value calculation unit 19. It is explanatory drawing for demonstrating the setting direction of a tuning gain. It is explanatory drawing for demonstrating the setting method of correction control command value K3 * C. It is explanatory drawing for demonstrating operation | movement of the vehicle body vibration control apparatus of 1st Embodiment. It is explanatory drawing for demonstrating the application example of the vehicle body vibration control apparatus of 1st Embodiment. It is a conceptual diagram of the structure of the vehicle of 2nd Embodiment. It is a block diagram showing the structure of the program which a microprocessor performs. 3 is a block diagram illustrating a configuration of a driving power wheel system vibration control unit 16. FIG. 3 is a flowchart showing the operation of a driving power wheel system vibration control unit 16. It is a figure for demonstrating operation | movement of a regulator gain and a tuning gain multiplication part.

Next, an embodiment according to the present invention will be described with reference to the drawings.
(First embodiment)
The vehicle body vibration control device of this embodiment is mounted on a front-wheel drive type four-wheel electric vehicle, and controls the sprung behavior of the vehicle body by controlling the torque generated by a braking / driving motor as a power source. . Specifically, the vehicle body vibration control device of the present embodiment is for enabling suppression of wheel load fluctuation, improvement of steering response, and suppression of roll behavior.

(Constitution)
FIG. 1 is a conceptual diagram illustrating a schematic configuration of a vehicle according to the first embodiment.
As shown in FIG. 1, the vehicle 1 includes a steering angle sensor 2, an accelerator opening sensor 3, a brake pedal depression force sensor 4, and a wheel speed sensor 5.
The steering angle sensor 2 is disposed on the steering column and detects the steering angle δo by the steering wheel 6. Then, the steering angle sensor 2 outputs a detection signal representing the detection result of the steering angle δo by the steering wheel 6 to the braking / driving motor ECU 12 described later.

The accelerator opening sensor 3 is disposed on the accelerator pedal and detects the accelerator opening. The accelerator opening is the amount of depression of the accelerator pedal. Then, the accelerator opening sensor 3 outputs a detection signal representing the detection result of the accelerator opening to the braking / driving motor ECU 12.
The brake pedal depression force sensor 4 is disposed on the brake pedal and detects the depression force of the brake pedal. Then, the brake pedal depression force sensor 4 outputs a detection signal representing the detection result of the depression force of the brake pedal to the braking / driving motor ECU 12.

The wheel speed sensor 5 is disposed on each of the wheels 5FL to 5RR, and detects the wheel speed VwFL to VwRR of each of the wheels 5FL to 5RR. The wheel speed sensor 5 outputs detection signals representing the wheel speeds VwFL to VwRR of the wheels 5FL to 5RR to the braking / driving motor ECU 12.
The vehicle 1 includes an inverter 7, a braking / driving motor 8, and a transmission 9. Here, the inverter 7, the braking / driving motor 8, and the transmission 9 constitute a torque adding means 100 described later.

The inverter 7 supplies the electric power stored in the battery 10 to the braking / driving motor 8 in accordance with the command output by the braking / driving motor ECU 12. Supply of power to the braking / driving motor 8 is performed by DC-AC conversion of the power of the battery 10 and AC current obtained by the conversion.
The braking / driving motor 8 generates torque according to the electric power supplied from the inverter 7. Then, the braking / driving motor 8 outputs the generated torque to the transmission 9.

The transmission 9 is disposed on a drive shaft 11 provided on each of the front wheels (drive wheels) 5FL and 5FR, and applies torque output from the braking / driving motor 8 to the front wheels 5FL and 5FR.
Further, the vehicle 1 includes a braking / driving motor ECU 12. Here, the braking / driving motor ECU 12 constitutes behavior estimation means 101, behavior estimation step, suppression torque control means 102, suppression torque control step, promotion torque control means 103, and promotion torque control step, which will be described later.

  The braking / driving motor ECU 12 includes a microprocessor. The microprocessor includes an integrated circuit including an A / D conversion circuit, a D / A conversion circuit, a central processing unit, a memory, and the like. Then, the braking / driving motor ECU 12 calculates the torque to be output to the braking / driving motor 8 based on the detection signals output from the sensors 2 to 5 according to the program stored in the memory, and issues a command to output the calculated torque to the inverter 7. Output to.

FIG. 2 is a block diagram illustrating a functional configuration of the vehicle according to the first embodiment.
As shown in FIG. 2, this block diagram includes torque adding means 100, behavior estimating means 101, suppression torque control means 102, and facilitating torque control means 103.
Torque adding means 100 adds driving torque to the wheels.
The behavior estimating means 101 estimates the sprung behavior of the vehicle body based on the driver's requested driving torque, the road surface disturbance applied from the road surface to the wheel, and the turning resistance applied to the wheel by steering. Then, the behavior estimation unit 101 outputs the estimation result of the sprung behavior of the vehicle body to the suppression torque control unit 102 and the promotion torque control unit 103.

The suppression torque control means 102 adds torque in a direction to suppress the fluctuation of the component caused by the required drive torque and the fluctuation of the component caused by the road surface disturbance among the components constituting the sprung behavior of the vehicle body estimated by the behavior estimation means 101. The driving torque applied by the means 100 is controlled.
The facilitating torque control means 103 is added by the torque adding means 100 in the direction of facilitating the fluctuation of the physical quantity related to the front wheel load, which is the component caused by the turning resistance among the components constituting the sprung behavior of the vehicle body estimated by the behavior estimating means 101. The driving torque to be controlled is controlled.

FIG. 3 is a block diagram showing a configuration of a program executed by the microprocessor.
The braking / driving motor ECU 12 constitutes the control block of FIG. 3 by a program executed by the microprocessor. This control block includes a driver request torque calculation unit 13, an adder 14, a torque command value calculation unit 15, and a driving power vehicle system vibration control unit 16. Here, the driver request torque calculator 13 constitutes a behavior estimation unit 101 and a behavior estimation step. Further, the adder 14 constitutes a suppression torque control means 102, a suppression torque control step, a promotion torque control means 102, and a promotion torque control step. Furthermore, the driving power wheel system vibration control unit 16 includes a behavior estimation unit 101, a behavior estimation step, a suppression torque control unit 102, a suppression torque control step, a promotion torque control unit 103, and a promotion torque control step.

  The driver request torque calculator 13 calculates the driver request torque based on the detection signal output from the accelerator opening sensor 3 and the detection signal output from the brake pedal depression force sensor 4. The driver request torque is an output torque requested by the driver to the braking / driving motor 8. The driver request torque is represented by a motor end value that is a torque value on the rotating shaft of the braking / driving motor 8. Then, the driver request torque calculation unit 13 outputs the calculated driver request torque to the adder 14 and the driving power vehicle system vibration control unit 16.

  In the present embodiment, an example in which the driver request torque is calculated based on the detection signal output from the accelerator opening sensor 3 and the detection signal output from the brake pedal depression force sensor 4 has been described, but other configurations are employed. You can also. For example, it is good also as a structure which makes the detected value itself by various sensors, such as the accelerator opening sensor 3, etc. into a driver request torque.

  The adder 14 adds the driver torque correction value output from the driving force vehicle system vibration control unit 16 to the driver request torque output from the driver request torque calculation unit 13. As a result, the driver request torque is corrected. The driver torque correction value is a correction value calculated by the driving force vehicle system vibration control unit 16 as described later based on the driver required torque, the wheel speeds VwFL to VwRR, and the steering angle δo. Then, the driver request torque calculation unit 13 outputs the corrected driver request torque to the torque command value calculation unit 15 as the corrected request torque.

  The torque command value calculation unit 15 outputs to the braking / driving motor 8 based on the corrected required torque output from the adder 14 and the output of other systems such as VDC (Vehicle Dynamics Control) and TCS (Traction Control System). Calculate the torque to be applied. Then, the torque command value calculator 15 outputs the calculated torque to the inverter 7 as a torque command value.

FIG. 4 is a block diagram illustrating the configuration of the driving power wheel system vibration control unit 16.
FIG. 5 is a flowchart showing the operation of the driving power wheel system vibration control unit 16.
As shown in FIG. 4, the driving power vehicle vibration control unit 16 includes an input conversion unit 17, a vehicle body vibration estimation unit 18, and a torque command value calculation unit 19. Here, the input conversion part 17 comprises the behavior estimation means 101 and a behavior estimation step. Moreover, the vehicle body vibration estimation part 18 comprises the behavior estimation means 101 and a behavior estimation step. Further, the torque command value calculation unit 19 constitutes an assist torque control means 103 and a torque control step.

  The input conversion unit 17 inputs information of the detection signals output from the steering angle sensor 2, the accelerator opening sensor 3, the brake pedal depression force sensor 4, and the wheel speed sensor 5 to the vehicle model 26 used by the vehicle body vibration estimation unit 18. Convert to format. Specifically, the input conversion unit 17 includes a drive torque conversion unit 20, a suspension stroke calculation unit 21, a vertical force conversion unit 22, a vehicle body speed estimation unit 23, a turning behavior estimation unit 24, and a turning resistance estimation unit 25. Here, the vertical force converting unit 22 and the turning resistance estimating unit 25 constitute a behavior estimating unit 101 and a behavior estimating step.

  The drive torque converter 20 reads the driver request torque output by the driver request torque calculator 13 (step S101 in FIG. 5). Subsequently, the drive torque converter 20 multiplies the read driver request torque by the gear ratio of the transmission 9. Thus, the driver request torque is converted from the motor end value to the drive shaft end value (step S102 in FIG. 5). Here, the drive shaft end value is a torque value at the front wheels 5FL and 5FR. Then, the drive torque conversion unit 20 outputs the multiplication result as the drive torque Tw to the vehicle body vibration estimation unit 18. Here, the drive torque Tw is a drive shaft end value of the required drive torque of the dry request torque.

FIG. 6 is a block diagram showing the configuration of the suspension stroke calculation unit 21.
FIG. 7 is a flowchart showing the operation of the suspension stroke calculation unit 21.
The suspension stroke calculation unit 21 determines the suspension stroke amounts Zf and Zr and the stroke speed dZf of the front and rear wheels 5FL to 5RR based on the detection signal output from the wheel speed sensor 5, that is, the detection signal indicating the wheel speeds VwFL to VwRR. dZr is calculated. Specifically, as shown in FIG. 6, the suspension stroke calculation unit 21 includes an average front wheel speed calculation unit 34, an average rear wheel speed calculation unit 35, a front wheel band pass filter processing unit 36, and a rear wheel band pass filter process. Unit 37, front wheel suspension stroke calculation unit 38, and rear wheel suspension stroke calculation unit 39.

  The average front wheel speed calculation unit 34 reads the detection signals output from the wheel speed sensors 5 of the front wheels 5FL and 5FR (step S103 in FIG. 5 and step S201 in FIG. 7). Subsequently, the average front wheel speed calculation unit 34 calculates the average front wheel speed VwF = (VwFL + VwFR) / 2 based on the read detection signal (step S202 in FIG. 7). Then, the average front wheel speed calculation unit 34 outputs the calculated average front wheel speed VwF to the front wheel band-pass filter processing unit 36.

  The average rear wheel speed calculation unit 35 reads the detection signals output from the wheel speed sensors 5 of the rear wheels 5RL and 5RR (step S103 in FIG. 5 and step S201 in FIG. 7). Subsequently, the average rear wheel speed calculation unit 35 calculates an average rear wheel speed VwR = (VwRL + VwRR) / 2 based on the read rear wheel speeds VwRL and VwRR (step S202 in FIG. 7). Then, the average rear wheel speed calculation unit 35 outputs the calculated average rear wheel speed VwR to the rear wheel band-pass filter processing unit 37.

  The front wheel band pass filter processing unit 36 extracts only components near the vehicle body resonance frequency from the average front wheel speed VwF output by the average front wheel speed calculation unit 34. Then, the front wheel band pass filter processing unit 36 outputs the extracted vehicle body resonance frequency vicinity vibration component fVwF to the front wheel suspension stroke calculation unit 38 and the rear wheel suspension stroke calculation unit 39 (step S203 in FIG. 7).

The rear wheel band-pass filter processing unit 37 extracts only components near the vehicle body resonance frequency from the average rear wheel speed VwR output by the average rear wheel speed calculation unit 35. Then, the rear wheel band pass filter processing unit 37 outputs the extracted vehicle body resonance frequency vicinity vibration component fVwR to the front wheel suspension stroke calculation unit 38 and the rear wheel suspension stroke calculation unit 39 (step S203 in FIG. 7).
Thus, in this embodiment, only the components fVwF and fVwR in the vicinity of the vehicle body resonance frequency are extracted from the average front wheel speed VwF and the average rear wheel speed VwR. Therefore, wheel speed fluctuations and noise components due to acceleration / deceleration of the entire vehicle 1 can be removed from the average front wheel speed VwF and the average rear wheel speed VwR, and only the wheel speed component representing vehicle body vibration can be extracted.

FIG. 8 is a diagram for explaining a method of calculating the stroke amount of the suspension.
FIG. 9 is a graph showing the front wheel suspension geometry characteristics.
FIG. 10 is a graph showing the rear wheel suspension geometry characteristics.
The front wheel suspension stroke calculation unit 38 calculates the longitudinal displacement Xtf of the front wheels 5FL, 5FR based on the vehicle body resonance frequency vicinity vibration component fVwF extracted by the front wheel band pass filter processing unit 36. Subsequently, the front wheel suspension stroke calculation unit 38 performs time differentiation on the calculated longitudinal displacement Xtf to calculate a time differential value dXtf. Subsequently, the front wheel suspension stroke calculation unit 38 calculates a suspension stroke amount Zf and a stroke speed dZf according to the following equations (1) and (2) based on the calculated longitudinal displacement Xtf and time differential value dXtf (FIG. 7). Step S204). Then, the front wheel suspension stroke calculation unit 38 outputs the calculation result to the vertical force conversion unit 22.
Zf = KgeoF · Xtf (1)
dZf = KgeoF · dXtf (2)

  Here, KgeoF is the gradient near the origin of the graph representing the front wheel suspension geometry characteristic of FIG. In the graph of FIG. 9, the horizontal axis represents the longitudinal displacement Xtf of the front wheels 5FL, 5FR, the vertical axis represents the vertical displacement Zf of the vehicle body above the front wheels 5FL, 5FR, the longitudinal displacement Xtf, and the vertical displacement Zf of the vehicle body. It is a graph showing the relationship.

The rear wheel suspension stroke calculation unit 39 calculates the longitudinal displacement Xtr of the rear wheels 5RL and 5RR based on the vehicle body resonance frequency vicinity vibration component fVwR extracted by the rear wheel band pass filter processing unit 37. Subsequently, the rear wheel suspension stroke calculation unit 39 performs time differentiation on the calculated longitudinal displacement Xtr to calculate a time differentiation value dXtr. Subsequently, the rear wheel suspension stroke calculation unit 39 calculates the suspension stroke amount Zr and the stroke speed dZr according to the following equations (3) and (4) based on the calculated longitudinal displacement Xtr and time differential value dXtr. Then, the front wheel suspension stroke calculation unit 38 outputs the calculation result to the vertical force conversion unit 22.
Zr = KgeoR · Xtr (3)
dZr = KgeoR · dXtr (4)

  Here, KgeoR is the gradient near the origin of the graph representing the rear wheel suspension geometry characteristics of FIG. In the graph of FIG. 10, the horizontal axis represents the longitudinal displacement Xtr of the rear wheels 5RL, 5RR, the vertical axis represents the vertical displacement Zr of the vehicle body above the rear wheels 5RL, 5RR, and the longitudinal displacement Xtr and the vertical displacement of the vehicle body. It is a graph showing the relationship with Zr.

  Returning to FIG. 4, the vertical force conversion unit 22 constituting the behavior estimation unit 101 multiplies the stroke amount Zf output from the suspension stroke calculation unit 21 by the spring constant Kf and the stroke speed dZf output from the suspension stroke calculation unit 21. Is multiplied by an attenuation coefficient Cf. Here, the spring constant Kf is a spring constant of the suspension of the front wheels 5FL and 5FR. The damping coefficient Cf is a damping coefficient of the suspension (shock absorber) of the front wheels 4RL and 5FR. Then, the vertical force conversion unit 22 outputs the sum of these multiplication results to the vehicle body vibration estimation unit 18 as the vertical force Fzf of the front wheels 5FL, 5FR (step S105 in FIG. 5). Here, the vertical force is an external force applied to the vehicle body by road surface disturbance applied to the wheels 5FL to 5RR from the road surface.

  The vertical force conversion unit 22 multiplies the stroke amount Zr output from the suspension stroke calculation unit 21 by the spring constant Kr, and also multiplies the stroke speed dZr output from the suspension stroke calculation unit 21 by the damping coefficient Cr. Here, the spring constant Kr is a spring constant of the suspension of the rear wheels 5RL and 5RR. The damping coefficient Cr is a damping coefficient of the suspension (shock absorber) of the rear wheels 4RL and 5RR. Then, the vertical force conversion unit 22 outputs the sum of these multiplication results to the vehicle body vibration estimation unit 18 as the vertical force Fzr of the rear wheels 5RL and 5RR (step S105 in FIG. 5).

  In the present embodiment, the suspension stroke amount and stroke speed are calculated based on the wheel speeds VWFL to VWRR, and the vertical forces Fzf and Fzr are calculated based on the calculated stroke amount and stroke speed. Other configurations can also be employed. For example, a stroke sensor that detects the stroke amount of the suspension may be provided, and the vertical forces Fzf and Fzr may be calculated based on the stroke amount detected by the stroke sensor and the time differential value of the detection result. Moreover, it is good also as a structure which uses the detection value itself by various sensors, such as a stroke sensor, as the vertical forces Fzf and Fzr.

The vehicle body speed estimation unit 23 reads the detection signal output from the wheel speed sensor 5 of the rear wheels 5RL and 5RR (driven wheels). Subsequently, the average rear wheel speed calculation unit 35 calculates the vehicle body speed V = (VwRL + VwRR) / 2 based on the read detection signal. Then, the vehicle body speed estimation unit 23 outputs the calculated vehicle body speed V to the turning behavior estimation unit 24.
The turning behavior estimation unit 24 calculates the yaw angular velocity γ and the vehicle slip angle βv according to the following equations (5) and (6) based on the detection signal output by the vehicle body speed estimation unit 23 and the detection signal output by the steering angle sensor 2. To do. Then, the turning behavior estimation unit 24 outputs the calculated yaw angular velocity γ and the vehicle body side slip angle βv to the turning resistance estimation unit 25. The turning behavior estimation unit 24 also outputs the steering angle δo represented by the detection signal of the steering angle sensor 2 together with these calculation results.

  Where δ is the tire turning angle calculated based on the steering angle δo, L is the wheel base, Lf is the distance from the center of gravity of the vehicle body to the front axle, Lr is the distance from the center of gravity of the vehicle body to the rear axle, and m is the vehicle weight. is there. Cpf is the tire cornering power of the front wheels 5FL and 5FR, and Cpr is the tire cornering power of the rear wheels 5RL and 5RR.

The turning resistance estimating unit 25 constituting the behavior estimating means 101 is based on the yaw angular velocity γ, the vehicle body side slip angle βv, and the steering angle δo output from the turning behavior estimating unit 24 according to the following equation (7), and the front wheels 5FL, 5FR The turning resistance Fcf is calculated. Here, the turning resistance Fcf is a resistance applied to the wheels 5FL to 5RR from the road surface by steering, and is a vehicle longitudinal component of a lateral force applied to the wheels 5FL to 5RR due to the occurrence of a slip angle. Then, the turning resistance estimation unit 25 outputs the calculated turning resistance Fcf to the vehicle model 26.
Fcf = βf · Fyf (7)
βf = βv + Lf · γ / V−δ
Fyf = βf · Cpf
Here, βf is the slip angle of the front wheels 5FL, 5FR, and Fyf is the cornering force of the front wheels 5FL, 5FR.

Further, the turning resistance estimation unit 25 calculates the turning resistance Fcr of the rear wheels 5RL and 5RR according to the following equation (8) based on the yaw angular velocity γ, the vehicle body side slip angle βv, and the steering angle δo output by the turning behavior estimation unit 24. calculate. Then, the turning resistance estimation unit 25 outputs the calculated turning resistance Fcr to the vehicle model 26.
Fcr = βr · Fyr (8)
βr = βv−Lr ・ γ / V
Fyr = βr · Cpr
Here, βr is the slip angle of the rear wheels 5FL, 5FR, and Fyr is the cornering force of the rear wheels 5FL, 5FR.

In the present embodiment, the example in which the turning resistances Fcf and Fcr are calculated based on the vehicle body speed V and the steering angle δo has been shown, but other configurations may be employed. For example, the detection values themselves by various sensors such as the steering angle sensor 2 may be used as the turning resistances Fcf and Fcr.
The vehicle body vibration estimation unit 18 constituting the behavior estimation unit 101 includes components constituting the sprung behavior of the vehicle body based on the driving torque Tw, the vertical forces Fzf and Fzr, and the turning resistances Fcf and Fcr output from the input conversion unit 17. calculate. Specifically, the vehicle body vibration estimation unit 18 includes a vehicle model 26. Here, the vehicle model 26 constitutes behavior estimation means 101 and a behavior estimation step.

FIG. 11 is a diagram for explaining the vehicle model 26.
The vehicle model 26 that constitutes the behavior estimation means 101 includes a component caused by the driving torque Tw, a component caused by the vertical forces Fzf and Fzr, and a turning resistance Fcf and Fcr among components constituting the sprung behavior of the vehicle body. The component to be calculated is calculated. That is, the sprung behavior of the vehicle body can be expressed by various physical quantities, and each of these various physical quantities includes various components, and the vehicle model 26 includes the above-described 3 of the various components. One component is calculated separately. Here, as the sprung behavior of the vehicle body, rotational motion around the pitch axis of the vehicle body and vertical motion in the bounce direction can be employed. As physical quantities representing the sprung behavior of the vehicle body, the bounce speed dZv, the bounce amount Zv, the pitch angular velocity dSp, and the pitch angle Sp of the vehicle body can be employed. These physical quantities dZv, Zv, dSp, Sp are parameters necessary for defining the front wheel load Wf and the rear wheel load Wr as shown in the following equations (9) and (10).
Wf = −2Kf (Zv + Lf · θp) −2Cf (dZv + Lf · dθp / dt) (9)
Wr = −2Kr (Zv + Lr · θp) −2Cr (dZv−Lr · dθp / dt) (10)

  Specifically, the vehicle model 26 is based on the drive torque Tw output from the drive torque converter 20, and among the components constituting the sprung behavior of the vehicle body, components dZv1, Zv1, dSp1, Sp1 due to the drive torque Tw. Is calculated. The components dZv1, Zv1, dSp1, and Sp1 resulting from the drive torque Tw are calculated according to the following equations (11) and (12) with Fzf, Fzr, Fcf, and Fcr set to “0” (step S112 in FIG. 5). Then, the vehicle model 26 outputs the calculated components dZv1, Zv1, dSp1, Sp1 to the torque command value calculation unit 19.

Here, as shown in FIG. 11, Ip is the moment of inertia around the pitch axis, hcg is the height of the center of gravity of the vehicle body, Rt is the height of the center of gravity of the wheel, and θp is the pitch angle.
Further, the vehicle model 26 is based on the vertical forces Fzf and Fzr output from the vertical force converter 22, and among the components constituting the sprung behavior of the vehicle body, the components dZv2, Zv2, dSp2, which are caused by the vertical forces Fzf and Fzr, Sp2 is calculated. The components dZv2, Zv2, dSp2, and Sp2 due to the vertical forces Fzf and Fzr are calculated according to the above equations (11) and (12) with Tw, Fcf, and Fcr set to “0” (step S112 in FIG. 5). Then, the vehicle model 26 outputs the calculated components dZv2, Zv2, dSp2, and Sp2 to the torque command value calculation unit 19.

  In the present embodiment, an example is shown in which the component caused by the driving torque Tw and the components caused by the vertical forces Fzf and Fzr are calculated from the components constituting the sprung behavior of the vehicle body. Can also be adopted. For example, at least one of the bounce speed dZv, the bounce amount Zv, the pitch angular speed dSp, and the pitch angle Sp of the vehicle body, or a component that constitutes a composite value thereof, a component caused by the driving torque Tw, and a vertical force Fzf , The component due to Fzr may be calculated. As the composite value, for example, a value obtained by multiplying each of the bounce speed dZv, the bounce amount Zv, the pitch angular speed dSp, and the pitch angle Sp by the coefficient and totaling the multiplication results can be employed.

  Further, the vehicle model 26 is based on the turning resistances Fcf and Fcr output from the turning resistance estimation unit 25, and among the components constituting the sprung behavior of the vehicle body, the components dZv3, Zv3, dSp3 due to the vertical forces Fzf and Fzr, Sp3 is calculated. Calculation of the components dZv3, Zv3, dSp3, Sp3 due to the turning resistances Fcf, Fcr is performed according to the above equations (11) and (12) with Tw, Fzf, Fzr being “0” (step S112 in FIG. 5). Subsequently, the vehicle model 26 calculates components dWf3, dWr3, dSF3, SF3 caused by the turning resistances Fcf, Fcr among the components constituting the sprung behavior of the vehicle body based on the calculated components dZv3, Zv3, dSp3, Sp3. calculate. dWf is the fluctuation speed of the front wheel load, dWr is the fluctuation speed of the rear wheel load, dSF is the fluctuation speed of the front-rear balance, and the SF front-back balance. The components dWf3, dWr3, dSF3, and SF3 resulting from the turning resistances Fcf and Fcr are calculated according to the above equations (9) and (10) (step S112 in FIG. 5). The vehicle model 26 then outputs the calculated components dWf3, dWr3, dSF3, SF3 to the torque command value calculation unit 19.

  In the present embodiment, an example is shown in which components due to the turning resistances Fcf and Fcr are calculated from among the components constituting the sprung behavior of the vehicle body, but other configurations may be employed. For example, at least one of the front wheel load Wf, the front wheel load fluctuation speed dWf, the pitch angular speed dSp, and the pitch angle Sp, or a component that constitutes a combined value thereof, a component caused by the turning resistances Fcf and Fcr, and It is good also as a structure which calculates the component resulting from the vertical forces Fzf and Fzr. As the composite value, for example, a value obtained by multiplying each of the front wheel load Wf, the front wheel load fluctuation speed dWf, the pitch angular speed dSp, and the pitch angle Sp by a coefficient and summing the multiplication results can be employed.

FIG. 12 is an explanatory diagram for explaining the operation of the torque command value calculation unit 19.
FIG. 13 is an explanatory diagram for explaining the setting direction of the tuning gain.
A torque command value calculation unit 19 that constitutes the assist torque control unit 103 calculates a driver torque correction value based on a component that constitutes the sprung behavior of the vehicle body output from the vehicle body vibration estimation unit 18. Specifically, the torque command value calculation unit 19 includes a first regulator 27, a second regulator 28, a third regulator 29, a first tuning gain multiplication unit 30, a second tuning gain multiplication unit 31, and a third tuning gain multiplication unit. 32, and a motor torque converter 33. Here, the 1st regulator 27 and the 2nd regulator 28 comprise the suppression torque control means 102 and the suppression torque control step. Moreover, the 3rd regulator 29 comprises the promotion torque control means 103 and the promotion torque control step. Further, the first tuning gain multiplication unit 30 and the second tuning gain multiplication unit 31 constitute a suppression torque control unit 102 and a suppression torque control step. Further, the third tuning gain multiplication unit 32 constitutes an assist torque control means 103 and an assist torque control step. Further, the motor torque conversion unit 33 constitutes a torque adding unit 100.

As shown in FIG. 12, the first regulator 27 uses components dZv1, Zv1, dSp1, and Sp1 resulting from the drive torque Tw output from the vehicle model 26 as state quantities x (= [dZv1, Zv1, dSp1, Sp1]). Multiply by regulator gain F1 and “−1”. Here, the regulator gain F1 is a gain for calculating a driving torque that converges the pitch angular velocity dSp1, which is a component caused by the driving torque Tw, to “0” by multiplying the state quantity x. For example, the regulator gain F1 is set according to the following equations (13) and (14).
F1 = R ⁻¹ B ^T P (13)

Here, the above equation (13) is a formula for calculating the regulator gain F1 of the optimum regulator that suppresses fluctuations in the pitch angular velocity dSp1, which is a component resulting from the drive torque Tw. In the above equation (14), J is a quadratic evaluation function in the optimal regulator, and P is a positive definite symmetric matrix that is a solution of the Riccati algebraic equation PA + A ^T P-PBR ^-1 B ^T P + Q = 0. Note that the following regulator gain F2 is also set according to a similar mathematical expression. As a result, the first regulator 27 corrects (controls) the drive torque in a direction that suppresses fluctuations in the pitch angular velocity dSp1, which is a component caused by the drive torque Tw, among the components constituting the sprung behavior of the vehicle body. Is calculated. Then, the first regulator 27 outputs the calculation result to the first tuning gain multiplication unit 30 as a drive torque control command value A.

  The second regulator 28 uses the components dZv2, Zv2, dSp2, and Sp2 resulting from the vertical forces Fzf and Fzr output from the vehicle model 26 as state quantities x (= [dZv2, Zv2, dSp2, Sp2]) and regulator gains F2 and “ -1 ". Here, the regulator gain F2 is a gain for calculating a driving torque that converges the pitch angle Sp2 that is a component caused by the vertical forces Fzf and Fzr to “0” by multiplying the state quantity x. As a result, the second regulator 28 corrects (controls) the drive torque in a direction that suppresses fluctuations in the pitch angle Sp2, which is a component caused by the vertical forces Fzf and Fzr, among the components constituting the sprung behavior of the vehicle body. A correction value is calculated. Then, the second regulator 28 outputs the calculation result to the second tuning gain multiplier 31 as a drive torque control command value B.

The third regulator 29 uses the components dWf3, dWr3, dSF3, and SF3 resulting from the turning resistances Fcf and Fcr output from the vehicle model 26 as state quantities Cx (= [dWf3, dWr3, dSF3, SF3]) and regulator gains F3 and “ -1 ". Here, the regulator gain F3 is a gain for calculating a driving torque for converging the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr, to “0” by multiplying the state quantity Cx. For example, the regulator gain F3 is set according to the following equations (15) and (16).
F3 = R ⁻¹ B ^T P (15)

  Here, the above equation (15) is a formula for calculating the regulator gain F3 of the optimum regulator that suppresses the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr. In the above equation (16), J is a quadratic evaluation function in the optimal regulator, and P is a positive definite symmetric matrix that is the solution of the Riccati algebraic equation. As a result, the third regulator 29 corrects the drive torque in a direction that suppresses fluctuations in the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr among the components constituting the sprung behavior of the vehicle body ( A correction value to be controlled) is calculated. Then, the third regulator 29 outputs the calculation result to the third tuning gain multiplier 32 as the drive torque control command value C.

  In the present embodiment, the example in which the control command value C is calculated based on the state quantity Cx, that is, the fluctuation speed dWf3 of the front wheel load is shown, but other configurations may be employed. For example, the control command value C may be calculated based on other physical quantities related to the front wheel load Wf, such as the front wheel load Wf and the fluctuation acceleration of the front wheel load. Further, the control command value C may be calculated based on physical quantities related to the pitch behavior of the vehicle body, such as the pitch angular velocity dSp and the pitch angle Sp.

The first tuning gain multiplier 30 multiplies the control command value A output from the first regulator 27 by the tuning gain K1 (step S113 in FIG. 5). Then, the first tuning gain multiplication unit 30 outputs the multiplication result to the motor torque conversion unit 33 as the modified control command value K1 · A. Here, the braking / driving motor ECU 12 adds the control command value A to the driver required torque, thereby changing the pitch angular velocity dSp1, which is a component caused by the drive torque Tw, among the components constituting the sprung behavior of the vehicle body. Can be suppressed. However, in the method of simply adding the drive torque control command value A to the driver request torque, the front and rear G may fluctuate and give the driver a sense of discomfort. Therefore, as shown in FIG. 13, the tuning gain K1 is set to a positive value and a value that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. As a result, the first tuning gain multiplication unit 30 prevents the driver from feeling uncomfortable due to the fluctuation of the front and rear G, and suppresses the fluctuation of the component caused by the driving torque Tw, that is, the fluctuation of the wheel load. The correction value of the driver request torque is adjusted in the direction to be adjusted.
As described above, in the present embodiment, the correction value of the driver request torque is adjusted so as to suppress the fluctuation of the pitch angular velocity dSp1, which is a component caused by the drive torque Tw. Therefore, fluctuations in wheel load can be suppressed and riding comfort can be improved.

Returning to FIG. 4, the second tuning gain multiplier 31 multiplies the control command value B output from the second regulator 28 by the tuning gain K2 (step S114 in FIG. 5). Then, the second tuning gain multiplication unit 31 outputs the multiplication result to the motor torque conversion unit 33 as the modified control command value K2 · B. Here, the braking / driving motor ECU 12 adds the control command value B to the driver request torque, and among the components constituting the sprung behavior of the vehicle body, the pitch angular velocity dSp2 that is a component caused by the vertical forces Fzf and Fzr Variations can be suppressed. However, in the method of simply adding the control command value B of the drive torque, the front and rear G may fluctuate and give the driver a sense of discomfort. Therefore, as shown in FIG. 13, the tuning gain K2 is set to a positive value and a value that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. Thereby, while preventing the driver from feeling uncomfortable due to fluctuations in the front-rear G, the second tuning gain multiplication unit 31 suppresses fluctuations in components caused by the vertical forces Fzf and Fzr, that is, fluctuations in wheel load. Adjust the correction value of the driver request torque.
As described above, in the present embodiment, the correction value of the driver request torque is adjusted in such a direction as to suppress the fluctuation of the pitch angular velocity dSp2 that is a component caused by the vertical forces Fzf and Fzr. Therefore, fluctuations in wheel load can be suppressed and riding comfort can be improved.

  Returning to FIG. 4, the third tuning gain multiplier 32 multiplies the control command value C output from the third regulator 29 by the tuning gain K3 (step S115 in FIG. 5). Then, the third tuning gain multiplication unit 32 outputs the multiplication result to the motor torque conversion unit 33 as the modified control command value K3 · C. Here, the braking / driving motor ECU 12 adds the control command value C to the driver request torque, and among the components constituting the sprung behavior of the vehicle body, fluctuations in the front wheel load that are components due to the turning resistances Fcf and Fcr. The fluctuation of the speed dWf3 can be suppressed. However, in the method of simply adding the control command value C of the drive torque, the front and rear G may fluctuate and give the driver a feeling of strangeness. Further, when the steering operation is started, by suppressing the fluctuation of the fluctuation speed dWf3 of the front wheel load, the nose dive behavior is suppressed, the increase of the front wheel load Wf is suppressed, and the cornering power Cp of the front wheels 5FL and 5FR is reduced. there's a possibility that. Therefore, the lateral force of the front wheels 5FL and 5FR is reduced, and the steering response may be reduced. Therefore, the tuning gain K3 is set to a negative value and a value that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. As a result, the third tuning gain multiplication unit 32 corrects the driver required torque in a direction that promotes fluctuations in components caused by the turning resistances Fcf and Fcr, while preventing the driver from feeling uncomfortable due to fluctuations in the longitudinal G. Adjust the value.

  As described above, in this embodiment, the correction value of the driver request torque is adjusted in a direction that promotes the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr. Therefore, when the steering operation is started, the driving torque can be reduced, the nose dive behavior can be promoted, and the front wheel load Wf can be increased. Thereby, the front wheel load Wf increases, the cornering power Cp of the front wheels 5FL, 5FR can be increased, the lateral force of the front wheels 5FL, 5FR can be increased, and the steering response can be improved. Further, by improving the steering response while suppressing the fluctuation of the wheel load, the linearity of the yaw angular velocity γ, that is, the linearity of the output with respect to the input can be improved. Thereby, the change of lateral G can be made loose and a roll behavior can be suppressed.

FIG. 14 is an explanatory diagram for explaining a method of setting the correction control command value K3 · C.
The roll behavior is generated by the lateral G accompanying the steering operation. That is, as shown in FIG. 14A, when the driver performs a steering operation, a lateral force is generated on the front wheels 5FL and 5FR. When a lateral force is generated on the front wheels 5FL and 5FR, a lateral G is generated on the vehicle body. When a lateral G is generated in the vehicle body, a roll angular velocity is generated. Therefore, the time waveform of the roll angular velocity has a correlation with the time waveform of the lateral G time differential value. Therefore, if the time differential value of the lateral G can be reduced, the absolute value of the peak value of the roll angular velocity can be reduced. At that time, if the temporal differential value of the lateral G is simply reduced, the lateral G is reduced and the yaw angular velocity is reduced, so that the steering response is reduced. Therefore, if the time differential value of the lateral G is reduced and the timing at which the lateral G starts to increase, that is, the rising of the lateral G can be accelerated, the steering responsiveness is reduced while reducing the absolute value of the peak value of the roll angular velocity. Can be improved. Then, when the driver starts the steering operation, such a lateral G indicates the ideal torque for reducing the drive torque before the roll angular velocity generated by the steering operation reaches the peak after the steering operation is started. This can be realized by adding to 5RR. In this manner, the driving torque is reduced, so that the front wheel load Wf is increased and the cornering power Cp of the front wheels 5FL and 5FR is increased. Therefore, the response of the yaw angular velocity γ increases, and the lateral force applied to the front wheels 5FL, 5FR increases at an early timing. Further, the lateral differential value of the lateral G can be reduced by increasing the lateral force at an early timing.

  Here, as shown in FIG. 14B, when a lateral force is generated by the steering operation, a turning resistance is generated along with the lateral G. Therefore, the time waveform of the turning resistance has a correlation with the time waveform of the lateral G. Further, when a turning resistance is generated, a pitch angular velocity is generated. Therefore, the time waveform of the pitch angular velocity has a correlation with the time waveform of the differential value of the turning resistance. Therefore, the time waveform of the roll angular velocity resulting from the steering operation can be predicted based on the time waveform of the pitch angular velocity resulting from the turning resistance, that is, the time waveform of the pitch behavior. Focusing on the fluctuation component of the front wheel load that can be predicted from the pitch behavior, that is, the fluctuation speed dWf of the front wheel load, when the driver starts the steering operation, the roll angular velocity peaks due to the steering operation after the steering operation is started. Before reaching the peak, the peak waveform reaches a peak. This is the same characteristic as the time waveform of the ideal torque by changing the positive / negative sign to a negative value. Therefore, the state quantity Cx including the fluctuation speed dWf of the front wheel load is multiplied by the regulator gain F3 and “−1”, and the multiplication result is multiplied by the regulator gain K1 (<0) so as to suppress the fluctuation of the roll angular speed. A correction value for correcting (controlling) the driving torque can be calculated.

  The motor torque conversion unit 33 multiplies the total value of the correction control command value K 1 · A, the correction control command value K 2 · B, and the correction control command value K 3 · C by the gear ratio of the transmission 9. As a result, the total value is converted from the drive shaft end value to the motor end value. Then, the motor torque converter 33 outputs the multiplication result as a driver torque correction value to the adder 14 (step S116 in FIG. 5).

(Operation)
FIG. 15 is an explanatory diagram for explaining the operation of the vehicle body vibration control apparatus of the first embodiment. FIG. 15A shows a time waveform of a physical quantity representing the operation of the first embodiment. FIG. 15B shows a time waveform of a physical quantity representing the operation of the comparative example.
Next, the operation of the vehicle 1 equipped with the vehicle body vibration control device will be described with reference to FIG.
First, while traveling on a highway, in order for the driver to drive the vehicle 1 straight ahead at a constant speed, the accelerator opening is kept constant, the steering wheel 6 is held at the origin position, and at time t0 in FIG. As shown, it is assumed that the steering input is set to “0”. Then, as shown in FIG. 3, the driver request torque calculation unit 13 of the braking / driving motor ECU 12 performs the driver request torque based on the detection signal output from the accelerator opening sensor 3 and the detection signal output from the brake pedal depression force sensor 4. Is calculated. Then, the driver request torque calculation unit 13 outputs the calculated driver request torque to the adder 14 and the input conversion unit 17. When the driver request torque calculation unit 13 outputs the driver request torque, as shown in FIG. 4, the drive torque conversion unit 20 of the input conversion unit 17 multiplies the driver request torque by the gear ratio of the transmission 9, and the multiplication result is obtained. The driving torque Tw is output to the vehicle model 26. When the drive torque converter 20 outputs the drive torque Tw, the vehicle model 26 calculates components dZv1, Zv1, dSp1, Sp1 due to the drive torque Tw among the components constituting the sprung behavior of the vehicle body. Then, the vehicle model 26 outputs the calculation results dZv1, Zv1, dSp1, Sp1 to the first regulator 27. When the vehicle model 26 outputs the components dZv1, Zv1, dSp1, Sp1, the first regulator 27 calculates the control command value A for the drive torque based on the components dZv1, Zv1, dSp1, Sp1, and the calculated control command value A Is output to the first tuning gain multiplier 30. Thus, the control command value A for correcting the driving torque in a direction in which the first regulator 27 suppresses the fluctuation of the pitch angular velocity dSp1, which is a component caused by the driving torque Tw, among the components constituting the sprung behavior of the vehicle body. calculate. When the first regulator 27 outputs the control command value A, the first tuning gain multiplication unit 30 multiplies the control command value A by the tuning gain K1, and uses the multiplication result as the corrected control command value K1 · A, and the motor torque conversion unit 33. Output to. As a result, the first tuning gain multiplication unit 30 makes the driver feel uncomfortable due to the fluctuation of the front and rear G while suppressing the fluctuation of the pitch angular velocity dSp1, which is a component caused by the driving torque Tw, that is, the fluctuation of the wheel load. The control command value A of the drive torque is adjusted in a direction to prevent the above.

  The suspension stroke calculation unit 21 calculates suspension stroke amounts Zf and Zr and stroke speeds dZf and dZr based on detection signals output from the wheel speed sensor 5, and outputs the calculation results to the vertical force conversion unit 22. When the input conversion unit 17 outputs the stroke amounts Zf and Zr and the stroke speeds dZf and dZr, the vertical force conversion unit 22 calculates the vertical forces Fzf and Fzr based on the stroke amount Zf and the stroke speed dZf, and the calculated vertical force Fzf and Fzr are output to the vehicle model 26. When the vertical force conversion unit 22 outputs the vertical forces Fzf and Fzr, the vehicle model 26 calculates components dZv2, Zv2, dSp2, and Sp2 due to the vertical forces Fzf and Fzr among the components constituting the sprung behavior of the vehicle body. Then, the calculation result is output to the second regulator 28. When the vehicle model 26 outputs the components dZv2, Zv2, dSp2, Sp2, the second regulator 28 calculates the control command value B of the drive torque based on the components dZv2, Zv2, dSp2, Sp2 caused by the vertical forces Fzf, Fzr. To do. Then, the second regulator 28 outputs the calculated control command value B to the second tuning gain multiplier 31. Thus, the control command value for correcting the driving torque in a direction in which the second regulator 28 suppresses fluctuations in the pitch angular velocity dSp2 that is a component caused by the vertical forces Fzf and Fzr among the components constituting the sprung behavior of the vehicle body. B is calculated. When the second regulator 28 outputs the control command value B, the second tuning gain multiplication unit 31 multiplies the control command value B by the tuning gain K2, and sets the multiplication result as the corrected control command value K2 · B, and the motor torque conversion unit 33. Output to. As a result, the second tuning gain multiplier 31 suppresses fluctuations in the pitch angular velocity dSp1, which is a component caused by the vertical forces Fzf and Fzr, that is, fluctuations in the wheel load, and makes the driver feel uncomfortable due to fluctuations in the front and rear G. The control command value B of the drive torque is adjusted in a direction to prevent the application.

  The vehicle body speed estimation unit 23 calculates the vehicle body speed V based on the detection signal output from the wheel speed sensor 5, and outputs the calculated vehicle body speed V to the turning behavior estimation unit 24. When the vehicle body speed estimation unit 23 outputs the vehicle body speed V, the turning behavior estimation unit 24 calculates the yaw angular velocity γ and the vehicle body side slip angle βv based on the vehicle body speed V and the detection signals output from the steering angle sensor 2, and the yaw angular velocity. The γ and the vehicle body side slip angle βv are output to the turning resistance estimation unit 25. Here, the calculation results of the yaw angular velocity γ and the vehicle body side slip angle βv are “0” because the steering input is “0”. When the turning behavior estimating unit 24 outputs the yaw angular velocity γ and the vehicle body side slip angle βv, the turning resistance estimating unit 25 turns the turning resistances Fcf, Fcr (= 0) based on the yaw angular velocity γ, the vehicle body side slip angle βv, and the tire turning angle δ. ) And the calculated turning resistances Fcf and Fcr are output to the vehicle model. When the turning resistance estimation unit 25 outputs the turning resistances Fcf and Fcr, the vehicle model 26 includes components dWf3, dWr3, dSF3, SF3 (= 0) and the calculation result is output to the third regulator 29. When the vehicle model 26 outputs the components dWf3, dWr3, dSF3, SF3, the third regulator 29 controls the drive torque control command value C (=) based on the components dWf3, dWr3, dSF3, SF3 caused by the turning resistances Fcf, Fcr. 0) and the calculated control command value C is output to the third tuning gain multiplier 32. When the third regulator 29 outputs the control command value C, the third tuning gain multiplication unit 32 multiplies the control command value C by the tuning gain K3, and sets the multiplication result as the corrected control command value K3 · C (= 0). Output to the torque converter 33.

  Then, the motor torque converter 33 multiplies the total value of the correction control command value K1 · A, the correction control command value K2 · B, and the correction control command value K3 · C by the gear ratio of the transmission 9, and the multiplication result is obtained. It outputs to the adder 14 as a driver torque correction value. When the motor torque converter 33 outputs the driver torque correction value, as shown in FIG. 3, the adder 14 corrects the driver request torque by adding the driver torque correction value to the driver request torque, and corrects the correction result. The torque is then output to the torque command value calculator 15 as the post request torque. When the adder 14 outputs the corrected required torque, the torque command value calculation unit 15 calculates a torque command value based on the corrected required torque, and outputs the calculated torque command value to the inverter 7. When the torque command value calculation unit 15 outputs the torque command value, the inverter 7 supplies the electric power stored in the battery 10 to the braking / driving motor 8 according to the output torque command value. The braking / driving motor 8 generates torque according to the electric power supplied from the inverter 7 and applies the generated torque to the front wheels 5FL and 5RR via the transmission 9 and the drive shaft 11. As a result, as shown at times t0 to t1 in FIG. 15A, the driving torque of the front wheels 5FL and 5RR is controlled, and fluctuations in the pitch angular velocity dSp can be suppressed as compared with the case where the driving torque is not corrected. Therefore, fluctuations in wheel load, that is, vibration of the vehicle body can be suppressed, and riding comfort can be improved.

  Here, it is assumed that the driver starts the steering operation by the steering wheel 6 in order to change the lane of the vehicle 1 and gradually increases the steering input as shown at time t1 in FIG. Then, as the absolute value of the steering input gradually increases, the turning behavior estimation unit 24 calculates values with gradually increasing absolute values as the yaw angular velocity γ and the vehicle body side slip angle βv. Further, the turning resistance estimation unit 25 calculates values with gradually increasing absolute values as the turning resistances Fcf and Fcr. Further, the vehicle model 26 gradually calculates a value having a large absolute value as the components dWf3, dWr3, dSF3, SF3 caused by the turning resistances Fcf, Fcr. Then, the third regulator 29 corrects the driving torque in a direction to suppress the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr among the components constituting the sprung behavior of the vehicle body. A command value C (> 0) is calculated. That is, at the start of the steering operation, nose dive behavior occurs due to the turning resistances Fcf and Fcr, and the fluctuation speed dWf3 of the front wheel load increases, so that the drive torque is increased in a direction to suppress the increase of the fluctuation speed dWf3. It is assumed that the drive torque is corrected in the direction to be generated. When the third regulator 29 outputs the control command value C, the third tuning gain multiplier 32 multiplies the control command value C by a negative tuning gain K3 (<0), and the multiplication result is corrected to the corrected control command value K3 · C (<0) is output to the motor torque converter 33. As a result, the third tuning gain multiplying unit 32 increases the front wheel load fluctuation speed dWf3 to increase the driving torque as a direction to promote the fluctuation of the front wheel load fluctuation speed dWf3, which is a component caused by the turning resistances Fcf and Fcr. The control command value C is adjusted in the direction in which the value decreases.

  Therefore, the torque command value calculation unit 15 gradually reduces the torque command value through the motor torque conversion unit 33 and the adder 14. Then, the braking / driving motor 8 gradually reduces the generated torque through the inverter 7. As a result, when the nose down behavior occurs as the sprung behavior of the vehicle body, the nose down behavior can be promoted by reducing the drive torque as shown at time t1 in FIG. Therefore, the front wheel load Wf can be increased and the cornering power Cp of the front wheels 5FL and 5FR can be increased as compared with the case where the drive torque is not corrected. Therefore, the response of the yaw angular velocity γ can be improved, the lateral force acting on the front wheels 5FL and 5FR can be increased, and the steering response of the vehicle 1 can be improved. As a result, the responsiveness of the yaw angular velocity γ increases, and the lateral force applied to the front wheels 5FL, 5FR increases at an early timing. Further, the lateral differential value of the lateral G is reduced by increasing the lateral force at an early timing. Here, the time waveform of the roll angular velocity has a correlation with the time waveform of the lateral G time differential value. Therefore, the absolute value of the peak value of the roll angular velocity can be reduced by reducing the time differential value of the lateral G.

Further, by improving the steering response while suppressing the fluctuation of the wheel load, the linearity of the yaw angular velocity γ, that is, the linearity of the output with respect to the input can be improved. Thereby, the change of lateral G can be made loose and the fluctuation | variation of a roll angular velocity can also be suppressed.
Furthermore, it is assumed that the yaw angular velocity γ is converging to a constant value after the steering operation by the steering wheel 6 is started. Then, by reducing the time differential value of the yaw angular velocity γ, the so-called entrainment phenomenon in which the yaw angular velocity γ increases and the movement of the vehicle 1 increases can be suppressed.

  As shown in FIG. 15B, the method of correcting the driving torque only in the direction of suppressing the wheel load fluctuation can suppress the vibration of the vehicle body, but causes a decrease in steering response and an increase in roll behavior. In addition, the driver may feel uncomfortable. For example, when the driver starts the steering operation, the turning resistances Fcf and Fcr applied to the wheels 5FL to 5RR increase, and the driving force decreases. Therefore, when nose-down behavior occurs, correction is made in a direction that suppresses the nose-down behavior, that is, in a direction that increases the drive torque. Therefore, the front wheel load Wf is reduced, and the cornering power Cp of the front wheels 5FL, 5FR is reduced. As a result, the steering response decreases, the response of the yaw angular velocity γ decreases, the lateral force applied to the front wheels 5FL, 5FR decreases, and the steering response of the vehicle 1 decreases. In addition, since the steering response decreases, the linearity of the yaw angular velocity γ, that is, the linearity of the output with respect to the input decreases. As a result, the change in the lateral G cannot be moderated, and the roll behavior increases. Further, when the yaw angular velocity γ converges, the yaw angular velocity γ increases and the entrainment phenomenon is worsened.

  As described above, in the present embodiment, the inverter 7, the braking / driving motor 8, the transmission 9, and the motor torque conversion unit 33 in FIG. 4 constitute the torque adding means 100. In the same manner, the braking / driving motor ECU 12 in FIG. 1, the driver request torque calculation unit 13 in FIG. 3, the driving force vehicle system vibration control unit 16, the input conversion unit 17, the vehicle body vibration estimation unit 18, the torque command in FIGS. The value calculation unit 19 constitutes behavior estimation means 101 and a behavior estimation step. Further, the vertical force conversion unit 22, the turning resistance estimation unit 25 in FIG. 4 and the vehicle model 26 in FIG. 4 also constitute the behavior estimation unit 101 and the behavior estimation step. Further, the braking / driving motor ECU 12 of FIG. 1, the adder 14 of FIG. 3, the driving power vehicle system vibration control unit 16, and the torque command value calculation unit 19 constitute a suppression torque control means 103 and a suppression torque control step. Further, the first regulator 27, the second regulator 28, the first tuning gain multiplication unit 30, and the second tuning gain multiplication unit 31 also constitute the suppression torque control means 103 and the suppression torque control step. Further, the braking / driving motor ECU 12 of FIG. 1, the adder 14 of FIG. 3, the driving force vehicle system vibration control unit 16, the torque command value calculation unit 19, the third regulator 29, and the third tuning gain multiplication unit 32 are promoted torque control means. 103 and the assist torque control step.

(Effect of this embodiment)
(1) In the present embodiment, the suppression torque control means 102 controls the drive torque in a direction that suppresses fluctuations in the component caused by the required drive torque and road surface disturbance among the components constituting the sprung behavior of the vehicle body. Further, the promotion torque control means 103 controls the drive torque in a direction that promotes fluctuations in the front wheel load, which is a component caused by turning resistance.
According to such a configuration, before starting the steering operation, the wheel load fluctuation is suppressed by controlling the driving torque in a direction to suppress the fluctuation of the component caused by the driver's requested driving torque and road surface disturbance. can do. Also, when the steering operation is started, by controlling the driving torque in the direction that promotes the fluctuation of the front wheel load caused by the turning resistance applied to the wheel, the nose down behavior can be promoted, the wheel load of the front wheel can be increased, the steering Responsiveness can be improved. Further, by improving the steering response while suppressing the fluctuation of the wheel load, the change in the lateral G can be moderated, and the roll behavior can be suppressed when the steering operation is performed.

(2) The behavior estimation unit 101 estimates rotational motion about the pitch axis of the vehicle body and vertical motion in the bounce direction as the sprung behavior of the vehicle body based on the required drive torque, road disturbance, and turning resistance. Further, the promotion torque control means 103 controls roll torque that suppresses the roll behavior of the vehicle body by controlling the drive torque in the direction that promotes the fluctuation of the physical quantity related to the front wheel load, which is a component caused by the turning resistance, based on the estimation result. Take control.
According to such a configuration, the physical quantity that can be controlled by the drive torque can be estimated as the sprung behavior of the vehicle body. Therefore, the roll behavior can be appropriately suppressed by controlling the driving torque.

(3) The assist torque control means 103 reduces the drive torque as roll suppression control after the driver starts the steering operation and before the roll angular velocity generated by the steering operation reaches a peak.
According to such a configuration, by reducing the drive torque before the roll angular velocity reaches the peak, the nose-down behavior can be promoted, the wheel load of the front wheels can be increased, and the steering response can be improved. Further, by improving the steering response, that is, the yaw response, the timing at which the yaw angular velocity starts to increase can be advanced, and the subsequent fluctuations in the yaw angular velocity can be moderated. Therefore, the lateral G variation can be suppressed and the roll angular velocity variation can be suppressed.

(4) When the roll suppression control is performed, the assist torque control unit 103 performs the roll suppression control that controls the drive torque based on the physical quantity related to the pitch behavior of the vehicle body or the physical quantity related to the front wheel load.
According to such a configuration, the roll behavior due to steering is generated in substantially the same phase as the pitch behavior due to turning resistance. Therefore, by controlling the drive torque based on fluctuations in the pitch angular velocity of the vehicle body or fluctuations in the front wheel load, The driving torque can be reduced before the angular velocity reaches a peak.

(5) The facilitating torque control means 103 calculates the required driving torque and the driving torque that suppresses fluctuations in components caused by road surface disturbance, and the driving torque that is reduced by the torque adding means 100 by roll suppression control. Based on the total value, the drive torque is controlled.
According to such a configuration, by suppressing the vibration of the vehicle body due to the required drive torque and road surface disturbance, it is possible to more effectively suppress the roll behavior when the steering operation is performed.

(6) The behavior estimation unit 101 estimates the turning resistance based on the steering angle and the vehicle body speed.
According to such a configuration, the turning resistance can be calculated with a relatively easy configuration.
(7) The behavior estimation unit 101 estimates a road surface disturbance based on the wheel speed.
According to such a configuration, the road surface disturbance can be calculated with a relatively easy configuration.
(8) The behavior estimation means 101 estimates the stroke speed and stroke amount of the suspension based on the wheel speed, multiplies the estimated stroke speed and the suspension damping coefficient, and multiplies the estimated stroke amount by the spring constant of the suspension. Multiply And the behavior estimation means 101 makes the total value of these multiplication results the estimated value of a road surface disturbance.
According to such a configuration, the road surface disturbance can be calculated with an easier configuration.

(9) In the present embodiment, the suppression torque control step controls the drive torque in a direction that suppresses fluctuations in the component caused by the required drive torque and road surface disturbance among the components constituting the sprung behavior of the vehicle body. Further, in the assist torque control step, the drive torque is controlled in a direction that promotes fluctuations in the front wheel load, which is a component caused by turning resistance.
According to such a configuration, before starting the steering operation, the wheel load fluctuation is suppressed by controlling the driving torque in a direction to suppress the fluctuation of the component caused by the driver's requested driving torque and road surface disturbance. can do. Also, when the steering operation is started, by controlling the driving torque in the direction that promotes the fluctuation of the front wheel load caused by the turning resistance applied to the wheel, the nose down behavior can be promoted, the wheel load of the front wheel can be increased, the steering Responsiveness can be improved. Further, by improving the steering response while suppressing the fluctuation of the wheel load, the change in the lateral G can be moderated, and the roll behavior can be suppressed when the steering operation is performed.

(Application examples)
FIG. 16 is an explanatory diagram for explaining an application example of the vehicle body vibration control device of the first embodiment.
In the present embodiment, the braking / driving ECU 12 is driven in a direction that promotes the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr among the components constituting the sprung behavior of the vehicle body. Although an example of correcting the torque has been shown, other configurations can be employed. For example, as shown in FIG. 16, the driving torque is corrected in a direction that promotes the fluctuation of the fluctuation speed dWf3 of the front wheel load, and the components of the rear wheel load Wr3 and the rear wheel load, which are components caused by the turning resistances Fcf and Fcr, are corrected. The driving torque may be corrected in a direction to suppress the fluctuation speed dWr3. Alternatively, the driving torque may be corrected in a direction that suppresses at least one of the physical quantities related to the rear wheel load, such as the rear wheel load Wr3, the rear wheel load fluctuation speed dWr3, and the rear wheel load fluctuation acceleration.
(Effects of this application example)
(1) In this application example, the physical quantity related to the rear wheel load is estimated as a component caused by the turning resistance, and the drive torque is controlled in a direction to suppress the fluctuation of the estimated physical quantity related to the rear wheel load.
According to such a configuration, among the components caused by the turning resistance, particularly the component that is desired to promote the fluctuation (for example, any one of the front wheel load, the fluctuation speed of the front wheel load, the pitch angular speed, and the pitch angle, or a composite value thereof) ) Can concentrate the control effect.

(Second Embodiment)
Next, a second embodiment of the present invention will be described with reference to the drawings. In addition, the same code | symbol is used about the same structure as said each embodiment.
This embodiment is mounted on a four-wheel internal combustion engine vehicle (that is, an FR / MT vehicle) that is a rear wheel drive type and a manual transmission type, and controls the torque generated by the engine that is a power source, thereby The point of controlling the sprung behavior is different from the first embodiment.

FIG. 17 is a conceptual diagram of a configuration of a vehicle according to the second embodiment.
Specifically, as shown in FIG. 17, in this embodiment, instead of the inverter 7, the braking / driving motor 8, the transmission 9, the battery 10, the drive shaft 11, and the braking / driving motor ECU, the engine 50, the MT transmission 51 and ECM 52. Here, the engine 50 and the MT transmission 51 constitute the torque adding means 100. Further, the ECM 52 constitutes a behavior estimation means 101, a behavior estimation step, a suppression torque control means 102, a suppression torque control step, an assist torque control means 103, and an assist torque control step.

Engine 50 generates torque in accordance with a command output from ECM 52 and outputs the generated torque to MT transmission 51.
The MT transmission 51 applies torque output from the engine 50 to the rear wheels 5RL and 5RR via the shaft 53, the differential gear 54, and the drive shaft 55.
The ECM 52 includes a microprocessor. The microprocessor includes an integrated circuit including an A / D conversion circuit, a D / A conversion circuit, a central processing unit, a memory, and the like. Then, the ECM 52 calculates torque to be output to the engine 50 based on the detection signals output from the sensors 2 to 5 in accordance with a program stored in the memory, and outputs a command to output the calculated torque to the engine 50.

FIG. 18 is a block diagram showing a configuration of a program executed by the microprocessor.
Further, as shown in FIG. 18, the present embodiment is different from the first embodiment in that the torque command value calculation unit 15 outputs a torque command value to the engine 50.
FIG. 19 is a block diagram illustrating a configuration of the driving power wheel system vibration control unit 16.
FIG. 20 is a flowchart showing the operation of the driving power wheel system vibration control unit 16.

  As shown in FIG. 19, in the present embodiment, the drive torque converter 20, the first regulator 27, the second regulator 28, the first tuning gain multiplier 30, the second tuning gain multiplier 31, and the third tuning gain multiplier. The operation of the unit 32 is different from that of the first embodiment. Further, the present embodiment is different from the first embodiment in that an engine torque converter 56 is provided instead of the motor torque converter 33.

  The drive torque conversion unit 20 reads the driver request torque output by the driver request torque calculation unit 13 (step S101 in FIG. 20). The driver request torque is represented by an engine end value that is a torque value on the rotation shaft of the engine 50. Subsequently, the drive torque converter 20 multiplies the read driver request torque by a gear ratio. Here, the gear ratio is a ratio between the average rotational speed of the left and right rear wheels 5RL and 5RR that are drive wheels and the rotational speed of the engine 50. Thus, the driver request torque is converted from the engine end value to the drive shaft end value (step S102 in FIG. 20). The drive shaft end value is a torque value at the rear wheels 5RL and 5RR. Then, the drive torque conversion unit 20 outputs the multiplication result as the drive torque Tw to the vehicle body vibration estimation unit 18.

FIG. 21 is a diagram for explaining the operation of the regulator gain and tuning gain multiplication unit.
The first regulator 27 uses the components dZv1, Zv1, dSp1, and Sp1 resulting from the drive torque Tw output from the vehicle model 26 as state quantities x (= [dZv1, Zv1, dSp1, Sp1]), and regulator gains F1 and “−1. ”. Here, the regulator gain F1 is a gain for calculating a driving torque that converges the pitch angular velocity dSp1, which is a component caused by the driving torque Tw, to “0” by multiplying the state quantity x. For example, the regulator gain F1 is set according to the following equations (17) and (18).
F1 = R ⁻¹ B ^T P (17)

Here, the above equation (17) is a formula for calculating the regulator gain F1 of the optimum regulator that suppresses the fluctuation of the pitch angular velocity dSp1 among the components caused by the drive torque Tw. In the above equation (18), J is an evaluation function of a quadratic form in the optimal regulator, and P is a positive definite symmetric matrix that is a solution of the Riccati algebraic equation PA + A ^T P-PBR ⁻¹ B ^T P + Q = 0. Note that the following regulator gains F2, F4, and F5 are also set according to the same mathematical formula. As a result, the first regulator 27 corrects (controls) the drive torque in a direction that suppresses fluctuations in the pitch angular velocity dSp1, which is a component caused by the drive torque Tw, among the components constituting the sprung behavior of the vehicle body. Is calculated. Then, the first regulator 27 outputs the calculation result to the first tuning gain multiplication unit 30 as a drive torque control command value A.

  In addition, the first regulator 27 uses the components dZv1, Zv1, dSp1, and Sp1 resulting from the drive torque Tw output from the vehicle model 26 as the state quantity x (= [dZv1, Zv1, dSp1, Sp1]) as the regulator gain F4 and “ -1 ". Here, the regulator gain F4 is a gain for calculating the driving torque that converges the bounce speed dZv1 that is a component caused by the driving torque Tw to “0” by multiplying the state quantity x. As a result, the first regulator 27 corrects (controls) the drive torque in a direction that suppresses fluctuations in the bounce speed dZv1, which is a component caused by the drive torque Tw, among the components constituting the sprung behavior of the vehicle body. Is calculated. Then, the first regulator 27 outputs the calculation result to the first tuning gain multiplier 30 as a drive torque control command value D.

  The second regulator 28 uses the components dZv2, Zv2, dSp2, and Sp2 resulting from the vertical forces Fzf and Fzr output from the vehicle model 26 as state quantities x (= [dZv2, Zv2, dSp2, Sp2]) and regulator gains F2 and “ -1 ". Here, the regulator gain F2 is a gain for calculating a driving torque that converges the pitch angle Sp2 that is a component caused by the vertical forces Fzf and Fzr to “0” by multiplying the state quantity x. As a result, the second regulator 28 corrects (controls) the drive torque in a direction that suppresses fluctuations in the pitch angle Sp2, which is a component caused by the vertical forces Fzf and Fzr, among the components constituting the sprung behavior of the vehicle body. A correction value is calculated. Then, the second regulator 28 outputs the calculation result to the second tuning gain multiplier 31 as a drive torque control command value B.

  Further, the second regulator 28 uses the components dZv2, Zv2, dSp2, and Sp2 resulting from the vertical forces Fzf and Fzr output from the vehicle model 26 as the state quantity x (= [dZv2, Zv2, dSp2, Sp2]) and the regulator gain F5. And "-1". Here, the regulator gain F5 is a gain for calculating a driving torque for converging the bounce displacement Zv1, which is a component caused by the vertical forces Fzf and Fzr, to “0” by multiplying the state quantity x. As a result, the second regulator 28 corrects (controls) the drive torque in a direction that suppresses fluctuations in the bounce displacement Zv1, which is a component caused by the vertical forces Fzf and Fzr, among the components constituting the sprung behavior of the vehicle body. A correction value is calculated. Then, the second regulator 28 outputs the calculation result to the second tuning gain multiplier 31 as a drive torque control command value E.

The third regulator 29 uses the components dWf3, dWr3, dSF3, and SF3 resulting from the turning resistances Fcf and Fcr output from the vehicle model 26 as state quantities Cx (= [dWf3, dWr3, dSF3, SF3]) and regulator gains F3 and “ -1 ". Here, the regulator gain F3 is a gain for calculating a driving torque for converging the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr, to “0” by multiplying the state quantity Cx. For example, the regulator gain F3 is set according to the following equations (19) and (20).
F3 = R ⁻¹ B ^T P (19)

  Here, the above equation (19) is a formula for calculating the regulator gain F3 of the optimum regulator that suppresses the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr. In the above equation (20), J is an evaluation function of a quadratic form in the optimal regulator, and P is a positive definite symmetric matrix that is a solution of the Riccati algebraic equation. As a result, the third regulator 29 corrects the drive torque in a direction that suppresses fluctuations in the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr among the components constituting the sprung behavior of the vehicle body ( A correction value to be controlled) is calculated. Then, the third regulator 29 outputs the calculation result to the third tuning gain multiplier 32 as the drive torque control command value C.

  Further, the third regulator 28 uses the components dWf3, dWr3, dSF3, SF3 resulting from the turning resistances Fcf, Fcr output from the vehicle model 26 as the state quantity x (= [dWf3, dWr3, dSF3, SF3]) and the regulator gain F6. And "-1". Here, the regulator gain F6 is a gain for calculating a driving torque for converging the fluctuation speed dWr3 of the rear wheel load, which is a component caused by the turning resistances Fcf and Fcr, to “0” by multiplying the state quantity x. . As a result, the third regulator 29 corrects the drive torque in a direction that suppresses fluctuations in the fluctuation speed dWr3 of the rear wheel load, which is a component caused by the turning resistances Fcf and Fcr among the components constituting the sprung behavior of the vehicle body. A correction value to be controlled is calculated. Then, the third regulator 29 outputs the calculation result to the third tuning gain multiplier 32 as a drive torque control command value F.

The first tuning gain multiplication unit 30 multiplies the control command value A output from the first regulator 27 by the tuning gain K1 (step S301 in FIG. 20). Further, the first tuning gain multiplication unit 30 multiplies the control command value D output from the first regulator 27 by the tuning gain K4 (step S301 in FIG. 20). Then, the first tuning gain multiplication unit 30 outputs the multiplication result to the motor torque conversion unit 33 as the modified control command values K1 · A and K4 · D. Here, the ECM 52 adds the control command values A and D to the driver request torque, and among the components constituting the sprung behavior of the vehicle body, the pitch angular velocity dSp1 and the bounce velocity dZv1 are components caused by the drive torque Tw. Fluctuations can be suppressed. However, in the method of simply adding the drive torque control command values A and D, the front and rear G may fluctuate and give the driver a sense of discomfort. Therefore, as shown in FIG. 21, the tuning gains K1 and K4 are set to positive values and values in a range that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. Thereby, the first tuning gain multiplication unit 30 prevents the driver from feeling uncomfortable due to the fluctuation of the front and rear G, while suppressing the fluctuation of the component caused by the driving torque Tw, that is, the fluctuation of the wheel load. The correction value of the driver request torque is adjusted in the direction to be suppressed.
As described above, in the present embodiment, the correction value of the driver request torque is adjusted in a direction in which fluctuations in the pitch angular velocity dSp1 and the bounce velocity dZv1 that are components due to the drive torque Tw are suppressed. Therefore, fluctuations in wheel load can be suppressed and riding comfort can be improved.

The second tuning gain multiplication unit 31 multiplies the control command value B output from the second regulator 28 by the tuning gain K2 (step S302 in FIG. 20). Further, the second tuning gain multiplier 31 multiplies the control command value E output from the second regulator 28 by the tuning gain K5 (step S302 in FIG. 20). Then, the second tuning gain multiplication unit 31 outputs the multiplication result to the motor torque conversion unit 33 as the modified control command values K2 · B and K5 · E. Here, the ECM 52 adds the control command values B and E to the driver request torque, and among the components constituting the sprung behavior of the vehicle body, the pitch angular velocity dSp2 and the bounce that are components caused by the vertical forces Fzf and Fzr Variations in the amount Zv2 can be suppressed. However, in the method of simply adding the control command values B and E of the drive torque, the front and rear G may fluctuate and give the driver a sense of discomfort. Therefore, the tuning gains K2 and K5 are set to values that are positive values and do not give the driver a sense of incongruity due to fluctuations in the front and rear G. Thus, the second tuning gain multiplication unit 31 prevents the driver from feeling uncomfortable due to the fluctuation of the front and rear G, while suppressing the fluctuation of the component due to the vertical forces Fcf and Fcr, that is, the wheel load. The correction value of the driver request torque is adjusted in a direction to suppress the fluctuation.
As described above, in the present embodiment, the correction value of the driver request torque is adjusted in such a direction as to suppress the fluctuations in the pitch angular velocity dSp2 and the bounce amount Zv2, which are components caused by the vertical forces Fzf and Fzr. Therefore, fluctuations in wheel load can be suppressed and riding comfort can be improved.

  The third tuning gain multiplication unit 32 multiplies the control command value C output from the third regulator 29 by the tuning gain K3 (step S303 in FIG. 20). Further, the third tuning gain multiplication unit 32 multiplies the control command value F output from the third regulator 29 by the tuning gain K6 (step S303 in FIG. 20). Then, the third tuning gain multiplication unit 32 outputs the multiplication result to the motor torque conversion unit 33 as the corrected control command values K3 · C and K6 · F. Here, the ECM 52 adds the control command value C to the driver required torque, and among the components constituting the sprung behavior of the vehicle body, the fluctuation speed dWf3 of the front wheel load that is a component caused by the turning resistances Fcf and Fcr. Variations can be suppressed. However, in the method of simply adding the control command value C of the drive torque, the front and rear G may fluctuate and give the driver a feeling of strangeness. Further, when the steering operation is started, by suppressing the fluctuation of the fluctuation speed dWf3 of the front wheel load, the nose dive behavior is suppressed, the increase of the front wheel load Wf is suppressed, and the cornering power Cp of the front wheels 5FL and 5FR is reduced. there's a possibility that. Therefore, the lateral force of the front wheels 5FL and 5FR is reduced, and the steering response may be reduced. Therefore, the tuning gain K3 is set to a negative value and a value that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. Accordingly, the third tuning gain multiplication unit 32 promotes the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr, while preventing the driver from feeling uncomfortable due to the fluctuation of the front and rear G. Adjust the correction value of the driver request torque in the direction to start.

  As described above, in this embodiment, the correction value of the driver request torque is adjusted in a direction that promotes the fluctuation of the fluctuation speed dWf3 of the front wheel load, which is a component caused by the turning resistances Fcf and Fcr. Therefore, when the steering operation is started, the driving torque can be reduced, the nose dive behavior can be promoted, and the front wheel load Wf can be increased. Thereby, the front wheel load Wf increases, the cornering power Cp of the front wheels 5FL, 5FR can be increased, the lateral force of the front wheels 5FL, 5FR can be increased, and the steering response can be improved. Further, by improving the steering response while suppressing the fluctuation of the wheel load, the linearity of the yaw angular velocity γ, that is, the linearity of the output with respect to the input can be improved. Thereby, the change of lateral G can be made loose and a roll behavior can be suppressed.

  Further, in the present embodiment, by adding the corrected command value K3 · C to the driver request torque, when the steering operation is performed, the roll behavior is suppressed, more specifically, the peak of the roll angular velocity. The absolute value of the value can be reduced. However, the resonance frequency of the corrected command value K3 · C, that is, the resonance frequency of the pitch angular velocity dSp and the resonance frequency of the roll angular velocity are slightly different. Therefore, in the method of simply adding the corrected command value K3 · C, the phase of the corrected command value K3 · C, that is, the fluctuation speed dWf of the front wheel load caused by the turning resistances Fcf and Fcr is the peak value of the roll angular velocity. There is a possibility that the command value for reducing the absolute value is not optimal. Therefore, the third tuning gain multiplying unit 32 sets the value calculated based on the fluctuation speed dWr3 of the rear wheel load having a phase different from the fluctuation speed dWf of the front wheel load as a correction value for correcting the influence of the phase shift. . Therefore, the tuning gain K6 is a positive value and is set to a value that does not give the driver a sense of incongruity due to fluctuations in the front and rear G. Thereby, the command value for reducing the absolute value of the peak value of the roll angular velocity can be made more appropriate.

  In the present embodiment, the control command value F is calculated based on the fluctuation speed dWr3 of the rear wheel load, and the corrected command value K6 · F is calculated based on the calculated control command value F. Other configurations can also be employed. For example, the corrected command value K6 · F may be calculated based on other physical quantities related to the rear wheel load Wr, such as the rear wheel load Wr and the fluctuation acceleration of the rear wheel load.

The engine torque conversion unit 56 sets the gear ratio to the total value of the corrected control command values K1 · A, K4 · D, the corrected control command values K2 · B, K5 · E, and the corrected control command values K3 · C, K6 · F. Multiply. As a result, the total value is converted from the drive shaft end value to the engine end value. Then, the engine torque converter 56 outputs the multiplication result as a driver torque correction value to the adder 14 (step S304 in FIG. 20).
As described above, in the present embodiment, the engine 50 and the transmission 51 of FIG. Similarly, the ECM 52 in FIG. 17 constitutes a behavior estimation unit 101, a behavior estimation step, a suppression torque control unit 102, a suppression torque control step, an assist torque control unit 103, and an assist torque control step.

(Effect of this embodiment)
(1) In the present embodiment, the assist torque control means 103 corrects the driving torque to be reduced by the torque addition means 100 by roll suppression control based on the physical quantity related to the rear wheel load.
According to such a configuration, the resonance frequency of the pitch behavior of the vehicle body and the resonance frequency of the roll behavior are slightly different, and by correcting using physical quantities relating to the rear wheel load that is different in phase from the front wheel load, The driving torque can be made more appropriate.

7 is an inverter (torque adding means 100)
8 is a braking / driving motor (torque adding means 100).
9 is a transmission (torque adding means 100).
12 is a braking / driving motor ECU (behavior estimation means 101, behavior estimation step, suppression torque control means 102, suppression torque control step, assist torque control means 103, assist torque control step)
Reference numeral 13 denotes a driver request torque calculation unit (behavior estimation means 101, behavior estimation step).
14 is an adder (suppression torque control means 102, suppression torque control step, promotion torque control means 103, promotion torque control step)
Reference numeral 16 denotes a driving force vehicle vibration control unit (behavior estimation means 101, behavior estimation step, suppression torque control means 102, suppression torque control step, assist torque control means 103, assist torque control step).
17 is an input converter (behavior estimating means 101, behavior estimating step).
18 is a vehicle body vibration estimation unit (behavior estimation means 101, behavior estimation step).
19 is a torque command value calculation unit (behavior estimation means 101, behavior estimation step, suppression torque control means 102, suppression torque control step, assist torque control means 103, assist torque control step).
22 is a vertical force converter (behavior estimating means 101, behavior estimating step).
25 is a turning resistance estimation unit (behavior estimation means 101, behavior estimation step)
26 is a vehicle model (behavior estimation means 101, behavior estimation step).
Reference numeral 27 denotes a first regulator (suppression torque control means 102, suppression torque control step).
28 is a second regulator (suppression torque control means 102, suppression torque control step).
Reference numeral 29 denotes a third regulator (an assisting torque control means 103, an assisting torque control step).
Reference numeral 30 denotes a first tuning gain multiplier (suppression torque control means 102, suppression torque control step).
31 is a second tuning gain multiplier (suppression torque control means 102, suppression torque control step).
Reference numeral 32 denotes a third tuning gain multiplier (an assisting torque control means 103, an assisting torque control step).
Reference numeral 33 denotes a motor torque converter (torque adding means 100).
50 is an engine (torque adding means 100)
51 is an MT transmission (torque adding means 100).
52 is ECM (behavior estimation means 101, behavior estimation step, suppression torque control means 102, suppression torque control step, assist torque control means 103, assist torque control step)

Claims (8)

Torque adding means for adding driving torque to the wheels;
A behavior estimation means for estimating the sprung behavior of the vehicle body based on the driver's required drive torque, the road surface disturbance applied to the wheel from the road surface, and the turning resistance applied to the wheel by steering;
Of the components constituting the sprung behavior of the vehicle body estimated by the behavior estimating means, the torque adding means adds in a direction to suppress the fluctuation of the component caused by the required driving torque and the fluctuation of the component caused by the road surface disturbance. Suppression torque control means for controlling the driving torque to be
Controls the driving torque applied by the torque adding means in the direction that promotes the fluctuation of the physical quantity related to the front wheel load, which is the component caused by the turning resistance, among the components constituting the sprung behavior of the vehicle body estimated by the behavior estimating means. And an encouraging torque control means ,
The behavior estimating means estimates a rotational motion around the pitch axis of the vehicle body and a vertical motion in the bounce direction as the sprung behavior of the vehicle body based on the required drive torque, the road surface disturbance, and the turning resistance,
The facilitating torque control means is configured to change a physical quantity related to a front wheel load, which is a component caused by the turning resistance, based on the rotational motion around the pitch axis of the vehicle body estimated by the behavior estimating means and the vertical motion in the bounce direction. By controlling the driving torque applied by the torque adding means in the direction of promotion, roll suppression control for suppressing the roll behavior of the vehicle body is performed. Further, the promotion torque control means includes the roll suppression control. In the vehicle body vibration control device, the driving torque applied by the torque adding means is reduced after the driver starts the steering operation and before the roll angular velocity generated by the steering operation reaches a peak . The facilitator torque control means, when performing the rolling-restraining control, claim 1, characterized in that to reduce the driving torque is the torque addition means for adding on the basis of the physical quantity related to a physical quantity or front wheel load relating vehicle body pitching behavior The vehicle body vibration control device described in 1. The facilitator torque control means on the basis of the physical quantity related to the rear wheel load, the vehicle body vibration control apparatus according to claim 1 or 2, characterized in that to correct the driving torque to be reduced to the torque application means by the roll reduction control. The suppression torque control means calculates a drive torque that suppresses fluctuations in components caused by the required drive torque and the road surface disturbance,
The facilitating torque control means calculates a driving torque to be reduced by the torque adding means by the roll suppression control,
The torque adding means controls the driving torque added by the torque adding means based on a total value of the driving torque calculated by the suppression torque control means and the driving torque calculated by the promotion torque control means. The vehicle body vibration control device according to any one of claims 1 to 3 . The behavior estimation means, based on the steering angle and the vehicle speed, the vehicle body vibration control apparatus according to claim 1, any one of 4, characterized in that estimating the turning resistance. The behavior estimation means, based on the wheel speeds, the vehicle body vibration control apparatus according to any one of claims 1 5, characterized in that estimating the road surface disturbance. The behavior estimating means estimates a stroke speed and a stroke amount of the suspension based on a wheel speed, multiplies the estimated stroke speed and a suspension damping coefficient, and multiplies the estimated stroke amount and a spring constant of the suspension. The vehicle body vibration control device according to claim 6 , wherein a total value of these multiplication results is used as an estimated value of the road surface disturbance. A behavior estimation step for estimating the sprung behavior of the vehicle body based on the driver's required driving torque, the road surface disturbance applied to the wheel from the road surface, and the turning resistance applied to the wheel by steering;
Drive torque applied to the wheel in a direction that suppresses fluctuations in the component caused by the required drive torque and fluctuations in the component caused by the road surface disturbance among the components constituting the sprung behavior of the vehicle body estimated by the behavior estimation step Suppression torque control step for controlling
Assisting torque for controlling the driving torque applied to the wheel in the direction that promotes the fluctuation of the physical quantity related to the front wheel load, which is the component resulting from the turning resistance among the components constituting the sprung behavior of the vehicle body estimated by the behavior estimating step A control step ,
The behavior estimating step estimates a rotational motion around the pitch axis of the vehicle body and a vertical motion in the bounce direction as the sprung behavior of the vehicle body based on the required driving torque, the road surface disturbance, and the turning resistance,
The promoting torque control step is a step of calculating a change in a physical quantity related to a front wheel load, which is a component caused by the turning resistance, based on the rotational motion around the pitch axis of the vehicle body estimated by the behavior estimating means and the vertical motion in the bounce direction. By controlling the driving torque applied by the torque adding means in the direction of promotion, roll suppression control for suppressing the roll behavior of the vehicle body is performed, and the promotion torque control step includes the roll suppression control. As described above, after the driver starts the steering operation, the driving torque applied by the torque adding means is reduced before the roll angular velocity generated by the steering operation reaches the peak .

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