JP6078950B2 - Vehicle control device - Google Patents

Description

  The present invention relates to a control device that controls the state of a vehicle.

  As a technique related to a vehicle control device, a technique described in Patent Document 1 is disclosed. This publication discloses a technique for controlling the pitch motion of a vehicle using a plurality of acceleration sensors mounted on the vehicle.

JP 2009-18649 A

  However, if a plurality of acceleration sensors are used to control the pitch motion of the vehicle, the configuration becomes expensive, and it may be difficult to reduce the cost.

  The present invention has been made paying attention to the above problem, and an object of the present invention is to provide a vehicle control device capable of controlling the pitch motion of the vehicle with an inexpensive configuration.

In order to achieve the above object, in the vehicle control apparatus according to an embodiment of the present invention, the pitch rate suppression control amount is calculated based on the fluctuation of the wheel speed difference between the front and rear wheels to control the pitch motion.

  Therefore, pitch motion can be controlled with an inexpensive configuration without using a plurality of acceleration sensors.

1 is a system schematic diagram illustrating a vehicle control apparatus according to a first embodiment. FIG. 2 is a control block diagram illustrating a control configuration of the vehicle control device according to the first embodiment. It is a control block diagram showing the structure of the pitch rate suppression control for FF vehicles of Example 1. FIG. Represents the characteristic (f-S curve) between the tire slip ratio and the braking / driving force obtained by multiplying the friction coefficient by the ground contact load. 3 is a dead zone processing map used for pitch rate suppression control in the first embodiment. FIG. 3 is a control block diagram illustrating a configuration of pitch rate suppression control for the FR vehicle according to the first embodiment. FIG. 3 is a control block diagram illustrating a configuration of roll rate suppression control according to the first embodiment. 3 is a time chart illustrating an envelope waveform forming process of roll rate suppression control according to the first embodiment. FIG. 3 is a control block diagram illustrating a configuration of a traveling state estimation unit according to the first embodiment. It is a control block diagram showing the control content in the stroke speed calculating part of Example 1. FIG. 3 is a block diagram illustrating a configuration of a reference wheel speed calculation unit according to the first embodiment. It is the schematic showing a vehicle body vibration model. FIG. 6 is a control block diagram illustrating actuator control amount calculation processing when performing pitch control according to the first embodiment. It is a control block diagram showing brake pitch control of Example 1. It is the figure which expressed simultaneously the wheel speed frequency characteristic detected by the wheel speed sensor, and the stroke frequency characteristic of the stroke sensor which is not mounted in the Example. It is a control block diagram showing the frequency sensitive control in the sprung mass damping control of the first embodiment. It is a correlation diagram showing the human sensory characteristic in each frequency domain. It is a characteristic view showing the relationship between the vibration mixing ratio of the wing area | region by the frequency sensitive control of Example 1, and damping force. It is a figure showing the wheel speed frequency characteristic detected by the wheel speed sensor in a certain running condition. FIG. 3 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the first embodiment. FIG. 3 is a control block diagram illustrating a control configuration of a damping force control unit according to the first embodiment. 6 is a flowchart illustrating attenuation coefficient arbitration processing in a standard mode according to the first embodiment. 6 is a flowchart illustrating an attenuation coefficient arbitration process in the sport mode according to the first embodiment. 6 is a flowchart illustrating attenuation coefficient arbitration processing in the comfort mode according to the first embodiment. 6 is a flowchart illustrating attenuation coefficient arbitration processing in a highway mode according to the first exemplary embodiment. It is a time chart showing the attenuation coefficient change at the time of drive | working a wavy road surface and an uneven | corrugated road surface. 6 is a flowchart illustrating a mode selection process based on a running state in an attenuation coefficient arbitration unit according to the first embodiment.

[Example 1]
FIG. 1 is a system schematic diagram illustrating a vehicle control apparatus according to the first embodiment. The vehicle includes an engine 1 that is a power source and a brake 20 that generates braking torque due to friction force on each wheel (hereinafter, when displaying brakes corresponding to individual wheels, right front wheel brake: 20FR, left front wheel brake: 20FL). Right rear wheel brake: 20RR, left rear wheel brake: 20RL), and shock absorber 3 (hereinafter referred to as S / A) provided between each wheel and the vehicle body and capable of variably controlling the damping force. When displaying S / A corresponding to an individual wheel, right front wheel S / A: 3FR, left front wheel S / A: 3FL, right rear wheel S / A: 3RR, left rear wheel S / A: 3RL And).

  The engine 1 includes an engine controller (hereinafter also referred to as an engine control unit, which corresponds to a power source control means) 1a that controls torque output from the engine 1, and the engine controller 1a includes a throttle valve opening degree of the engine 1, By controlling the fuel injection amount, ignition timing, etc., the desired engine operating state (engine speed and engine output torque) is controlled. Further, the brake 20 generates a braking torque based on the hydraulic pressure supplied from the brake control unit 2 that can control the brake hydraulic pressure of each wheel according to the traveling state. The brake control unit 2 includes a brake controller (hereinafter also referred to as a brake control unit) 2a for controlling a braking torque generated by the brake 20, and a master cylinder pressure generated by a driver's brake pedal operation or a built-in motor. A pump pressure generated by the drive pump is used as a hydraulic pressure source, and a desired hydraulic pressure is generated in the brake 20 of each wheel by opening and closing operations of a plurality of solenoid valves.

  S / A3 is a damping force generator that attenuates the elastic motion of a coil spring provided between a vehicle unsprung (axle, wheel, etc.) and a sprung (vehicle body, etc.). It is configured to be variable. The S / A 3 includes a cylinder in which fluid is sealed, a piston that strokes in the cylinder, and an orifice that controls fluid movement between fluid chambers formed above and below the piston. Furthermore, orifices having a plurality of types of orifice diameters are formed in the piston, and an orifice corresponding to a control command is selected from the plurality of types of orifices when the S / A actuator is operated. Thereby, the damping force according to the orifice diameter can be generated. For example, if the orifice diameter is small, the movement of the piston is easily restricted, so that the damping force is high. If the orifice diameter is large, the movement of the piston is difficult to be restricted, and thus the damping force is small.

  In addition to the selection of the orifice diameter, for example, an electromagnetic control valve is arranged on the communication path connecting fluids formed above and below the piston, and the damping force is set by controlling the opening / closing amount of the electromagnetic control valve. There is no particular limitation. The S / A 3 has an S / A controller 3a (corresponding to damping force control means) that controls the damping force of the S / A 3, and controls the damping force by operating the orifice diameter by the S / A actuator.

  Further, a wheel speed sensor 5 for detecting the wheel speed of each wheel (hereinafter, when displaying the wheel speed corresponding to each wheel, right front wheel speed: 5FR, left front wheel speed: 5FL, right rear wheel speed: 5RR. , Left rear wheel speed: 5RL)), an integrated sensor 6 for detecting longitudinal acceleration, yaw rate and lateral acceleration acting on the center of gravity of the vehicle, and a steering angle which is a steering operation amount of the driver is detected. Steering angle sensor 7, vehicle speed sensor 8 for detecting vehicle speed, engine torque sensor 9 for detecting engine torque, engine speed sensor 10 for detecting engine speed, and master pressure sensor 11 for detecting master cylinder pressure. And a brake switch 12 that outputs an on-state signal when a brake pedal operation is performed, and an accelerator opening sensor 13 that detects an accelerator pedal opening. The signals of these various sensors are input to the S / A controller 3a. The arrangement of the integrated sensor 6 may be the position of the center of gravity of the vehicle, and may be a configuration that can estimate various values at the position of the center of gravity even at other locations, and is not particularly limited. Moreover, it is not necessary to be an integral type, and a configuration in which yaw rate, longitudinal acceleration, and lateral acceleration are individually detected may be employed.

  FIG. 2 is a control block diagram illustrating a control configuration of the vehicle control apparatus according to the first embodiment. In the first embodiment, the controller is composed of an engine controller 1a, a brake controller 2a, and an S / A controller 3a. Within the S / A controller 3a, there are a driver input control unit 31 for performing driver input control for achieving a desired vehicle posture based on driver operations (steering operation, accelerator operation, brake pedal operation, etc.), and various sensors. Based on the estimated traveling state, a traveling state estimating unit 32 that estimates the traveling state based on the detected value, a sprung mass damping control unit 33 that controls the vibration state on the spring based on the estimated traveling state, and the like. An unsprung vibration suppression control unit 34 that controls the unsprung vibration state, a shock absorber posture control amount output from the driver input control unit 31, and an unsprung vibration suppression control amount output from the sprung vibration suppression control unit 33. Based on the unsprung vibration suppression control amount output from the unsprung vibration suppression control unit 34, a damping force to be set in the S / A 3 is determined, and the damping force control unit 3 that performs the S / A damping force control. With the door.

  In the first embodiment, a configuration including three controllers as the controller is shown. However, for example, the damping force control unit 35 is excluded from the S / A controller 3a and used as an attitude control controller, and the damping force control unit 35 is configured as an S / A controller. The A controller may be configured to include four controllers, or each controller may be configured from one integrated controller without particular limitation. In the first embodiment, the engine controller and the brake controller in the existing vehicle are used as they are as the engine control unit 1a and the brake control unit 2a, and the S / A controller 3a is separately installed. It is assumed that the vehicle control apparatus according to the first embodiment is realized.

(Overall configuration of vehicle control device)
In the vehicle control apparatus of the first embodiment, three actuators are used to control the vibration state generated on the spring. At this time, since each control controls the sprung state, mutual interference becomes a problem. In addition, the elements that can be controlled by the engine 1, the elements that can be controlled by the brake 20, and the elements that can be controlled by the S / A 3 are different from each other, and how to control them in combination is a problem.
For example, the brake 20 can control a bounce motion and a pitch motion, but if both are performed, a feeling of deceleration is strong and a driver is likely to feel uncomfortable. In addition, S / A3 can control all of roll motion, bounce motion and pitch motion. However, when all control is performed by S / A3, the production cost of S / A3 is increased and the damping force is increased. Since it tends to be high, high-frequency vibration from the road surface side is likely to be input, and it is easy for the driver to feel uncomfortable. In other words, there is a trade-off that the control by the brake 20 does not cause deterioration of the high frequency vibration but increases the feeling of deceleration, and the control by the S / A3 does not cause the feeling of deceleration but invites the input of high frequency vibration. To do.

Therefore, in the vehicle control apparatus of the first embodiment, it is possible to comprehensively judge these problems and realize a control configuration that complements each other's weak points while taking advantage of the advantages as the respective control characteristics. Therefore, in order to realize a vehicle control apparatus that is inexpensive but has excellent vibration control capability, an overall control system was constructed mainly considering the points listed below.
(1) The control amount by the S / A 3 is suppressed by preferentially performing the control by the engine 1 and the brake 20.
(2) By limiting the control target motion of the brake 20 to the pitch motion, the feeling of deceleration in the control by the brake 20 is eliminated.
(3) The control amount by the engine 1 and the brake 20 is limited and output from the control amount that can be actually output, thereby reducing the burden on the S / A 3 and accompanying the control of the engine 1 and the brake 20. Suppresses discomfort that occurs.
(4) Skyhook control is performed by all actuators. At this time, without using a stroke sensor or a sprung vertical acceleration sensor generally required for skyhook control, the skyhook control can be performed with an inexpensive configuration using wheel speed sensors mounted on all vehicles. Realize.
(5) When performing sprung control by S / A3, scalar control (frequency sensitive control) is newly introduced for the input of high frequency vibration that is difficult to cope with by vector control such as skyhook control.
(6) By appropriately selecting the control state realized by the S / A 3 according to the traveling state, an appropriate control state according to the traveling state is provided.
The above is the outline of the entire control system configured in the embodiment. Hereinafter, individual contents for realizing these will be sequentially described.

(About the driver input controller)
First, the driver input control unit will be described. The driver input control unit 31 achieves the vehicle posture required by the driver by the engine side driver input control unit 31a that achieves the vehicle posture required by the driver by torque control of the engine 1 and the damping force control of S / A3. And an S / A side driver input control unit 31b. In the engine-side driver input control unit 31a, a pitch rate suppression control amount that suppresses the pitch motion of the vehicle that occurs during braking / driving, a ground load variation suppression control amount that suppresses ground load variation between the front wheels and the rear wheels, and the steering angle sensor 7 The yaw response control amount corresponding to the vehicle behavior that the driver wants to achieve is calculated based on the signal from the vehicle speed sensor 8 and output to the engine control unit 1a.
In the S / A side driver input control unit 31b, the driver wants to achieve based on the driver input damping force control amount that suppresses the pitch motion of the vehicle generated during braking / driving and the signals from the steering angle sensor 7 and the vehicle speed sensor 8. A driver input damping force control amount corresponding to the vehicle behavior is calculated and output to the damping force control unit 35. For example, when the driver is turning, if the nose side of the vehicle is lifted, the driver's field of view easily deviates from the road surface. In this case, the damping force of the four wheels is set so as to prevent the nose from rising by making the nose dive. Is output as a driver input damping force control amount. In addition, a driver input damping force control amount that suppresses the roll motion generated during turning is output.

(Pitch control by S / A side driver input control)
Here, the pitch rate suppression control performed by the S / A side driver input control unit 31b will be described. FIG. 3 is a control block diagram showing the configuration of pitch rate suppression control for an FF (front engine / front drive) vehicle according to the first embodiment.
In the S / A side driver input control unit 31b, a front / rear wheel speed difference calculation unit 312a that calculates a wheel speed difference between the front and rear wheels based on the value of each wheel speed sensor 5, and a road surface from the calculated front / rear wheel speed difference. The first road surface disturbance removing unit 312b that removes the front and rear wheel speed difference caused by the disturbance, and the front and rear slip that calculates the variation of the difference in the slip ratio (slip amount) of the front and rear wheels based on the front and rear wheel speed difference from which the road surface disturbance has been removed. Rate difference fluctuation calculation section 312c, second road disturbance removal section 312d that removes the front and rear slip ratio difference fluctuation caused by road disturbance from the calculated front and rear slip ratio difference fluctuation, and front and rear slip ratio difference fluctuation after removing the road disturbance And a driver input damping force control amount calculation unit 312e for calculating a damping force control amount for pitch rate suppression control for each of the front and rear wheels based on the minutes.

The front and rear wheel speed difference calculation unit 312a adds the wheel speed sensor values ωfr and ωfl detected by the wheel speed sensors 5 of the front wheels FR and FL, and divides this by 2 (ie, calculates the arithmetic mean). The front wheel speed ωf is calculated by removing the wheel speed difference caused by the roll motion. Similarly, the rear wheel speed ωr is calculated from the wheel speed sensor values ωrr and ωrl detected by the wheel speed sensor 5 of the rear wheels RR and RL. Then, the front-rear wheel speed difference Δω (= ωf−ωr) is calculated from the difference between the front wheel speed ωf and the rear wheel speed ωr.
The first road surface disturbance removal unit 312b performs band elimination filter processing on the calculated front / rear wheel speed difference Δω to remove the tire rotation primary vibration component, and performs low-pass filter processing to produce a sprung due to road surface disturbance. Remove front and rear wheel speed differences caused by behavioral changes (suspension stroke).

  The front / rear slip ratio difference fluctuation calculation unit 312c calculates the fluctuation dS of the front / rear wheel slip ratio difference ΔS by performing high-pass filter processing on the front / rear wheel speed difference Δω from which the road disturbance has been removed by the first road disturbance removal unit 312b. To do. That is, if Vb is the vehicle speed, the difference ΔS (= Sf−Sr) between the slip ratio Sf (= (Vb−ωf) / Vb) of the front wheels and the slip ratio Sr (= (Vb−ωr) / Vb) of the rear wheels ) Is −Δω / Vb. The high-pass filter process corresponds to time differentiation of the front / rear wheel speed difference Δω. Therefore, the fluctuation (time derivative) dω of the front and rear wheel speed difference Δω obtained by subjecting the front and rear wheel speed difference Δω to the high-pass filter processing is equivalent to the fluctuation (time differentiation) dS of the front and rear wheel slip ratio difference ΔS. . The front-rear wheel speed difference calculation unit 312a calculates the difference ΔS between the front and rear wheel slip ratios from the beginning, and the difference ΔS is time-differentiated by the front-rear slip ratio difference variation calculation unit 312c to calculate the variation dS. It is good as well.

The variation dS of the difference between the front and rear slip ratios has a correlation with the pitch motion. FIG. 4 shows a general characteristic (fS curve) between the tire slip ratio S and the braking / driving force f obtained by multiplying the friction coefficient μ by the ground load. When the contact load of the tire changes, the slip ratio of the tire that increases the contact load decreases, and the slip ratio of the tire that decreases the contact load increases. From the relationship of this change, it can be said that the fluctuation amount α of the slip ratio indicates the fluctuation amount of the ground contact load of the wheel. Therefore, the fluctuation amount of the slip ratio difference ΔS between the front and rear wheels can be handled as indicating the amount of movement of the ground load of the front and rear wheels (from the front wheel to the rear wheel or from the rear wheel to the front wheel), that is, the pitch rate.
Here, in the FF vehicle, the sign of the fluctuation dS of the slip ratio difference ΔS (= Sf−Sr) between the front and rear wheels differs between driving (acceleration) and braking (deceleration). Specifically, assuming that the slip ratio decreasing side is positive, the slip ratio of the front wheel (driving wheel) is smaller than that of the rear wheel (driven wheel) during driving, and the fluctuation dS of the slip ratio difference ΔS becomes positive. During braking, the slip ratio of the front wheel increases from the rear wheel, and the fluctuation dS of the slip ratio difference ΔS becomes negative.

  The second road surface disturbance removing unit 312d performs dead zone processing on the calculated fluctuation amount dS, and the front / rear slip ratio difference fluctuation caused by the minute road surface disturbance that cannot be removed by the low-pass filter processing of the first road surface disturbance removing unit 312b. Remove minutes. FIG. 5 is a vehicle speed sensitive map in which the dead zone threshold in the dead zone processing is changed according to the vehicle speed VSP. When the vehicle speed VSP is in the range from zero to the predetermined value VSP1, the threshold value is set to a predetermined constant value. Even in the region where the vehicle speed VSP is near zero (in this case, the absolute value of the fluctuation dS is very small), the damping force according to the threshold value is output, so that the control force is raised from the beginning of driving during driving and stopped during braking. It suppresses the swaying of time. In addition, since the higher the vehicle speed VSP, the more easily affected by the road surface disturbance, the fluctuation dS may be calculated smaller than the actual value. Therefore, in the range where the vehicle speed VSP is higher than the predetermined value VSP1, the width of the dead zone is widened by increasing the threshold according to the vehicle speed VSP. In this way, by increasing the width of the dead zone according to the vehicle speed VSP, the damping force according to the threshold value is secured regardless of the influence of the road surface disturbance. As a result, when the vehicle speed VSP is high, it only acts in a direction in which the pitch behavior during braking / driving is sufficiently suppressed, and there is no particular problem.

The driver input damping force control amount calculation unit 312e includes a front wheel side control amount calculation unit 312f and a rear wheel side control amount calculation unit 312g. The front wheel side control amount calculation unit 312f generates a front wheel side driver input damping force control amount Ff for pitch rate suppression control according to the front and rear slip ratio difference variation dS obtained by removing the road surface disturbance by the second road surface disturbance removing unit 312d. Calculate and output to the damping force control unit 35. Specifically, as shown in the front wheel side control amount calculation unit 312f in FIG. 3, in a region where dS is positive (during driving), dS is multiplied by a first gain to calculate a damping force F for driving. To do. During driving, the damping force on the extension side at S / A3 of the front wheel is increased in accordance with the increase in the absolute value of dS (increase in the amount of load movement to the rear wheel side) to suppress nose up or squat. In a region where dS is negative (during braking), dS is multiplied by the second gain to calculate damping force F for braking. During braking, the damping force on the contraction side in S / A3 of the front wheel is increased in accordance with the increase in the absolute value of dS (increase in the load movement amount to the front wheel side) to suppress the nose dive. Note that the second gain is set smaller than the first gain.
Similar to the front wheel side control amount calculation unit 312f, the rear wheel side control amount calculation unit 312g calculates a driver input damping force control amount Fr on the rear wheel side for pitch rate suppression control according to dS, and controls damping force. To the unit 35. During driving, the damping force on the contraction side in S / A3 of the rear wheel is increased in accordance with the increase in the absolute value of dS (increase in the amount of load movement to the rear wheel side) to suppress nose up or squat. During braking, the damping force on the extension side at S / A3 of the rear wheel is increased in accordance with the increase in the absolute value of dS (increase in the amount of load movement to the front wheel side), thereby suppressing nose diving.

FIG. 6 is a control block diagram showing the configuration of pitch rate suppression control for the FR (front engine / rear drive) vehicle of the first embodiment. Only parts different from the above-described control configuration for the front-wheel drive vehicle will be described.
The front / rear slip ratio difference variation calculation unit 312c includes a braking / driving determination unit 312h that determines whether braking or driving based on a detection value of the brake switch 12. The front / rear slip ratio difference variation calculation unit 312c is configured to determine whether the front / rear wheel slips based on the determination result of the braking / driving determination unit 312h and the variation | dS | of the front / rear wheel slip ratio difference ΔS | The variation dS of the rate difference ΔS is calculated.

  That is, in the FR vehicle, the sign of the fluctuation dS of the slip ratio difference ΔS (= Sf−Sr) between the front and rear wheels is the same during driving and braking. Specifically, when the slip rate decreasing side is positive, the slip ratio of the rear wheel (driving wheel) is reduced more than the front wheel (driven wheel) during driving, and the fluctuation dS of the slip ratio difference ΔS becomes negative. During braking, the slip ratio of the front wheel increases from the rear wheel, and the fluctuation dS of the slip ratio difference ΔS becomes negative. Therefore, as it is, it cannot be determined whether the variation dS is during driving or braking, and an appropriate control amount cannot be calculated. Therefore, for the FR vehicle, it is determined whether it is braking or driving, and a control amount is determined by attaching an appropriate sign to dS based on the determination result. Specifically, if the detected value of the brake switch 12 is OFF, it is determined that the vehicle is in driving (acceleration direction), and the sign of dS is set to be positive and output to the second road surface disturbance removing unit 312d. On the other hand, if the detected value of the brake switch 12 is ON, it is determined that braking is being performed (deceleration direction), and the sign of dS is set to be negative and is output to the second road surface disturbance removing unit 312d. Therefore, the driver input damping force control amount calculation unit 312e can set an appropriate damping force control amount at the time of braking and at the time of driving to realize pitch rate suppression control. The same applies to other rear wheel drive vehicles as well as FR vehicles. The determination of braking / driving may be performed by other means (for example, accelerator opening degree) regardless of the detection value of the brake switch 12.

As described above, in the first embodiment, the fluctuation dω (dS) of the detected wheel speed difference (slip rate difference) between the front and rear wheels is calculated, and the pitch rate suppression control amount is calculated based on the calculated fluctuation. Thus, the pitch motion can be controlled with an inexpensive configuration without using a plurality of acceleration sensors (for example, a vertical acceleration sensor provided corresponding to each wheel).
In addition, for example, it is conceivable to predict vehicle body fluctuations based on the accelerator opening and the brake pedal operation amount. In this case, however, it is necessary to set a vehicle model. On the other hand, in the first embodiment, the driver input damping force control amount F can be calculated simply by multiplying the variation dS by the gain. In other words, since the vehicle model is not used, the calculation load is small and the control device can be configured at low cost, and even if there is a specification change, there is no need to adjust the model for each vehicle. It can be adapted relatively easily to certain types of vehicles.
Further, the variation in the slip ratio of the tire has a close correlation with the variation in the contact load, while the variation in the contact load has a correlation with the variation on the spring (pitch motion) via the suspension. In the first embodiment, paying attention to this point, by calculating the pitch rate suppression control amount from the fluctuation of the slip ratio of the tire, with the same accuracy as the case of using the vertical acceleration sensor that detects the fluctuation on the spring, Pitch rate suppression control can be performed.
In addition to the driver input control on the S / A3 side, in the driver input control on the engine side, the pitch rate suppression control amount based on the fluctuation of the wheel speed difference (slip rate difference) between the front and rear wheels is set as the driving force control amount. ) It may be used. That is, any actuator may be used as long as it can suppress the pitch motion of the vehicle. Further, the above-described pitch rate suppression control using fluctuations in the wheel speed difference (slip rate difference) between the front and rear wheels is not limited to driver input control, but also other vehicle body attitude control, for example, sprung mass damping control (brake control, etc.) It is good also as applying to, and it does not specifically limit.

(About roll control by S / A side driver input control)
Next, roll rate suppression control performed by the S / A side driver input control unit 31b will be described. FIG. 7 is a control block diagram illustrating a configuration of roll rate suppression control according to the first embodiment. The lateral acceleration estimation unit 31b1 estimates the lateral acceleration Yg based on the front wheel steering angle δf detected by the steering angle sensor 7 and the vehicle speed VSP detected by the vehicle speed sensor 8. This lateral acceleration Yg is calculated from the following equation based on the vehicle body plan view model.
Yg = (VSP ² / (1 + A · VSP ² )) · δf
Here, A is a predetermined value.

  The 90 ° phase advance component creation unit 31b2 differentiates the estimated lateral acceleration Yg and outputs a lateral acceleration differential value dYg. The first addition unit 31b4 adds the lateral acceleration Yg and the lateral acceleration differential value dYg. The 90 ° phase delay component creation unit 31b3 outputs a component F (Yg) obtained by delaying the phase of the estimated lateral acceleration Yg by 90 °. The second adder 31b5 adds F (Yg) to the value added by the first adder 31b4. The Hilbert transform unit 31b6 calculates a scalar quantity based on the envelope waveform of the added value. The gain multiplication unit 31b7 multiplies the scalar amount based on the envelope waveform by the gain, calculates a driver input attitude control amount for roll rate suppression control, and outputs the calculated value to the damping force control unit 35.

  FIG. 8 is a time chart showing the envelope waveform forming process of the roll rate suppressing control according to the first embodiment. When the driver starts steering at time t1, roll rate begins to gradually occur. At this time, the 90 ° phase advance component is added to form an envelope waveform, and the driver input attitude control amount is calculated based on the scalar amount based on the envelope waveform, thereby suppressing the occurrence of roll rate in the initial stage of steering. it can. Next, when the driver enters the steering holding state at time t2, the 90 ° phase advance component disappears, and this time, the phase delay component F (Yg) is added. At this time, even if the roll rate itself does not change much in the steady turning state, a roll rate resonance component corresponding to the roll back is generated after the roll once. If the phase delay component F (Yg) is not added, the damping force from the time t2 to the time t3 is set to a small value, which may cause the vehicle behavior to become unstable due to the roll rate resonance component. In order to suppress this roll rate resonance component, a 90 ° phase delay component F (Yg) is added.

  When the driver shifts from the steered state to the straight traveling state at time t3, the lateral acceleration Yg decreases and the roll rate converges to a small value. Again, since the damping force is firmly secured by the action of the 90 ° phase delay component F (Yg), instability due to the roll rate resonance component can be avoided.

(About the running state estimation unit)
Next, the traveling state estimation unit will be described. FIG. 9 is a control block diagram illustrating the configuration of the traveling state estimation unit according to the first embodiment. In the traveling state estimation unit 32 of the first embodiment, basically, based on the wheel speed detected by the wheel speed sensor 5, the stroke speed of each wheel used for the skyhook control of the sprung mass damping control unit 33 described later, Calculate bounce rate, roll rate and pitch rate. First, the value of the wheel speed sensor 5 of each wheel is input to the stroke speed calculation unit 321, and the sprung speed is calculated from the stroke speed of each wheel calculated by the stroke speed calculation unit 321.

  FIG. 10 is a control block diagram illustrating control contents in the stroke speed calculation unit according to the first embodiment. The stroke speed calculation unit 321 is individually provided for each wheel, and the control block diagram shown in FIG. 10 is a control block diagram focusing on a certain wheel. In the stroke speed calculation unit 321, the value of the wheel speed sensor 5, the front wheel steering angle δf detected by the steering angle sensor 7, and the rear wheel steering angle δr (actual rear wheel steering if a rear wheel steering device is provided). The reference wheel speed calculation unit 300 that calculates a reference wheel speed based on the vehicle body lateral speed and the actual yaw rate detected by the integrated sensor 6, and the angle may be appropriately set to 0 in other cases. A tire rotation vibration frequency calculation unit 321a that calculates the tire rotation vibration frequency based on the calculated reference wheel speed, and a deviation calculation unit 321b that calculates a deviation (wheel speed fluctuation) between the reference wheel speed and the wheel speed sensor value. A GEO conversion unit 321c that converts the deviation calculated by the deviation calculation unit 321b into a suspension stroke amount, a stroke speed calibration unit 321d that calibrates the converted stroke amount to a stroke speed, A band elimination filter corresponding to the frequency calculated by the tire rotation vibration frequency calculation unit 321a is applied to the value calibrated by the roke speed calibration unit 321d to remove the tire rotation primary vibration component and calculate the final stroke speed. And a signal processing unit 321e.

[Regarding the reference wheel speed calculation unit]
Here, the reference wheel speed calculation unit 300 will be described. FIG. 11 is a block diagram illustrating a configuration of a reference wheel speed calculation unit according to the first embodiment. The reference wheel speed refers to a value obtained by removing various disturbances from each wheel speed. In other words, the difference between the wheel speed sensor value and the reference wheel speed is a value related to a component that fluctuates according to the stroke generated by the bounce behavior, roll behavior, pitch behavior, or unsprung vertical vibration of the vehicle body. In the example, the stroke speed is estimated based on this difference.

  In the plane motion component extraction unit 301, the first wheel speed V0 that is the reference wheel speed of each wheel is calculated based on the vehicle body plan view model with the wheel speed sensor value as an input. Here, the wheel speed sensor value detected by the wheel speed sensor 5 is ω (rad / s), the front wheel actual steering angle detected by the steering angle sensor 7 is δf (rad), and the rear wheel actual steering angle is δr (rad ), The vehicle body lateral speed is Vx, the yaw rate detected by the integrated sensor 6 is γ (rad / s), the vehicle speed estimated from the calculated reference wheel speed ω0 is V (m / s), and the reference to be calculated Wheel speed is VFL, VFR, VRL, VRR, front wheel tread is Tf, rear wheel tread is Tr, distance from vehicle center of gravity to front wheel is Lf, and distance from vehicle center of gravity to rear wheel is Lr. Using the above, the car body plan view model is expressed as follows.

(Formula 1)
VFL = (V−Tf / 2 ・ γ) cosδf + (Vx + Lf ・ γ) sinδf
VFR = (V + Tf / 2 ・ γ) cosδf + (Vx + Lf ・ γ) sinδf
VRL = (V−Tr / 2 ・ γ) cosδr + (Vx−Lr ・ γ) sinδr
VRR = (V + Tr / 2 ・ γ) cosδr + (Vx−Lr ・ γ) sinδr
If it is assumed that the vehicle is traveling normally without skidding, 0 may be input as the vehicle body lateral velocity Vx. When this is rewritten to a value based on V in each equation, it is expressed as follows. In this rewriting, V is described as V0FL, V0FR, V0RL, V0RR (corresponding to the first wheel speed) as a value corresponding to each wheel.
(Formula 2)
V0FL = {VFL−Lf · γsinδf} / cosδf + Tf / 2 · γ
V0FR = {VFR−Lf · γsinδf} / cosδf−Tf / 2 · γ
V0RL = {VRL + Lr · γsinδr} / cosδr + Tr / 2 · γ
V0RR = {VRR + Lf · γsinδf} / cosδr−Tr / 2 · γ

The roll disturbance removing unit 302 calculates the second wheel speeds V0F and V0R as the reference wheel speeds for the front and rear wheels based on the vehicle body front view model with the first wheel speed V0 as an input. The vehicle body front view model removes the wheel speed difference caused by the roll motion that occurs around the roll rotation center on the vertical line passing through the center of gravity of the vehicle when the vehicle is viewed from the front. Is done.
V0F = (V0FL + V0FR) / 2
V0R = (V0RL + V0RR) / 2
As a result, the second wheel speeds V0F and V0R from which disturbance based on the roll is removed are obtained.

The pitch disturbance removal unit 303 calculates the third wheel speeds VbFL, VbFR, VbRL, and VbRR, which are the reference wheel speeds for all the wheels, based on the vehicle side view model, with the second wheel speeds V0F and V0R as inputs. Here, the vehicle body side view model is to remove the wheel speed difference caused by the pitch motion generated around the pitch rotation center on the vertical line passing through the center of gravity of the vehicle when the vehicle is viewed from the lateral direction. It is expressed by the following formula.
(Formula 3)
VbFL = VbFR = VbRL = VbRR = {Lr / (Lf + Lr)} V0F + {Lf / (Lf + Lr)} V0R
In the reference wheel speed redistribution unit 304, VbFL (= VbFR = VbRL = VbRR) is substituted for V in the vehicle body plan view model shown in (Equation 1), and the final reference wheel speeds VFL, VFR, VRL of each wheel are respectively substituted. VRR is calculated and divided by the tire radius r0 to calculate the reference wheel speed ω0.

  When the reference wheel speed ω0 for each wheel is calculated by the above processing, a deviation between the reference wheel speed ω0 and the wheel speed sensor value is calculated, and this deviation is a wheel speed variation associated with the suspension stroke. Converted to stroke speed Vz_s. Basically, the suspension does not stroke only in the vertical direction when holding each wheel, but the wheel rotation center moves back and forth with the stroke, and the axle itself equipped with the wheel speed sensor 5 also tilts. It causes a rotation angle difference from the wheel. Since the wheel speed changes with this back-and-forth movement, the deviation between the reference wheel speed and the wheel speed sensor value can be extracted as the fluctuation accompanying this stroke. It should be noted that the degree of fluctuation may be set as appropriate according to the suspension geometry.

  When the stroke speed calculation unit 321 calculates the stroke speeds Vz_sFL, Vz_sFR, Vz_sRL, and Vz_sRR for each wheel by the above processing, the sprung speed calculation unit 322 calculates the bounce rate, roll rate, and pitch rate for skyhook control. Calculated.

(About the estimation model)
Skyhook control achieves a flat running state by controlling the posture of the sprung on the basis of the relationship between the stroke speed of S / A3 and the sprung speed. Here, in order to achieve the posture control on the spring by the skyhook control, it is necessary to feed back the sprung speed. Now, the value that can be detected from the wheel speed sensor 5 is the stroke speed, and since the vertical acceleration sensor or the like is not provided on the spring, the sprung speed needs to be estimated using an estimation model. Hereinafter, the problem of the estimation model and the model configuration to be adopted will be described.

  FIG. 12 is a schematic diagram showing a vehicle body vibration model. FIG. 12A shows a model of a vehicle (hereinafter referred to as a conveyor vehicle) having an S / A with a constant damping force, and FIG. 12B shows an S / A having a variable damping force. This is a model for performing skyhook control. In FIG. 12, Ms represents the mass on the spring, Mu represents the mass under the spring, Ks represents the elastic coefficient of the coil spring, Cs represents the damping coefficient of S / A, and Ku represents the unsprung (tire). , Cu represents an unsprung (tire) damping coefficient, and Cv represents a variable damping coefficient. Z2 represents a position on the spring, z1 represents a position under the spring, and z0 represents a road surface position.

When the convex vehicle model shown in FIG. 12 (a) is used, the equation of motion for the sprung is expressed as follows. Note that the first derivative (ie, speed) of z1 is represented by dz1, and the second derivative (ie, acceleration) is represented by ddz1.
(Estimation formula 1)
Ms · ddz2 = −Ks (z2−z1) −Cs (dz2−dz1)
When this relational expression is rearranged by Laplace transform, it is expressed as follows.
(Estimation formula 2)
dz2 = − (1 / Ms) · (1 / s ² ) · (Cs · s + Ks) (dz2−dz1)
Here, since dz2-dz1 is a stroke speed (Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR), the sprung speed can be calculated from the stroke speed. However, when the damping force is changed by the skyhook control, the estimation accuracy is remarkably lowered, and therefore, there is a problem that a large attitude control force (attenuating force change) cannot be given in the convex vehicle model.

Therefore, it is conceivable to use a vehicle model based on skyhook control as shown in FIG. Changing the damping force basically means changing the force that limits the piston moving speed of S / A 3 in accordance with the suspension stroke. Since the semi-active S / A3 that cannot positively move the piston in the desired direction is used, when the semi-active skyhook model is employed and the sprung speed is obtained, it is expressed as follows.
(Estimation formula 3)
dz2 = − (1 / Ms) · (1 / s ² ) · {(Cs + Cv) · s + Ks} (dz2−dz1)
However,
When dz2 · (dz2−dz1) ≧ 0 Cv = Csky · {dz2 / (dz2−dz1)}
When dz2 · (dz2−dz1) <0, Cv = 0
That is, Cv has a discontinuous value.

  Now, if you want to estimate the sprung speed using a simple filter, in the semi-active skyhook model, when this model is viewed as a filter, each variable corresponds to a filter coefficient, and the pseudo-differential term { Since (Cs + Cv) · s + Ks} includes a discontinuous variable attenuation coefficient Cv, the filter response becomes unstable and appropriate estimation accuracy cannot be obtained. In particular, when the filter response becomes unstable, the phase shifts. If the correspondence between the phase of the sprung speed and the sign is broken, the skyhook control cannot be achieved. Therefore, even when a semi-active S / A3 is used, it is not dependent on the sign relationship between the sprung speed and the stroke speed, and the sprung is performed using an active skyhook model that can directly use stable Csky. The speed was estimated. When the active sky hook model is adopted and the sprung speed is obtained, it is expressed as follows.

(Estimation formula 4)
dz2 =-(1 / s). {1 / (s + Csky / Ms)}. {(Cs / Ms) s + (Ks / Ms)} (dz2-dz1)
In this case, discontinuity does not occur in the pseudo differential term {(Cs / Ms) s + (Ks / Ms)}, and the {1 / (s + Csky / Ms)} term can be configured by a low-pass filter. Therefore, the filter response is stable and appropriate estimation accuracy can be obtained. Here, even if the active sky hook model is adopted, only semi-active control is actually possible, so the controllable area is halved. Therefore, the magnitude of the estimated sprung speed is smaller than the actual value in the frequency band below the sprung resonance, but the most important in skyhook control is the phase. If the correspondence between the phase and the sign can be maintained, the skyhook can be maintained. Since control is achieved and the magnitude of the sprung speed can be adjusted by other factors, there is no problem.

  From the above relationship, it can be understood that the sprung speed can be estimated if the stroke speed of each wheel is known. Next, since the actual vehicle is four wheels instead of one wheel, it is considered to estimate the state of the spring by mode decomposition into roll rate, pitch rate and bounce rate using the stroke speed of each wheel. To do. Now, when the above three components are calculated from the stroke speed of the four wheels, one corresponding component is insufficient, and the solution becomes indefinite. Therefore, a war plate representing the movement of the diagonal wheels is introduced. If the stroke amount bounce term is xsB, the roll term is xsR, the pitch term is xsP, the warp term is xsW, and the stroke amount corresponding to Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR is z_sFL, z_sFR, z_sRL, z_sRR, Holds.

以上の関係式から、xsB、xsR、xsP、xsWの微分ｄxsB等は以下の式で表される。
ｄxsB＝1/4（Vz_sFL＋Vz_sFR＋Vz_sRL＋Vz_sRR）
ｄxsR＝1/4（Vz_sFL−Vz_sFR＋Vz_sRL−Vz_sRR）
ｄxsP＝1/4（−Vz_sFL−Vz_sFR＋Vz_sRL＋Vz_sRR）
ｄxsW＝1/4（−Vz_sFL＋Vz_sFR＋Vz_sRL−Vz_sRR） (Formula 1)

From the above relational expression, the differential dxsB of xsB, xsR, xsP, xsW, etc. is expressed by the following expression.
dxsB = 1/4 (Vz_sFL + Vz_sFR + Vz_sRL + Vz_sRR)
dxsR = 1/4 (Vz_sFL−Vz_sFR + Vz_sRL−Vz_sRR)
dxsP = 1/4 (−Vz_sFL−Vz_sFR + Vz_sRL + Vz_sRR)
dxsW = 1/4 (−Vz_sFL + Vz_sFR + Vz_sRL−Vz_sRR)

Here, since the relationship between the sprung speed and the stroke speed is obtained from the estimation formula 4, among the estimation formula 4, − (1 / s) · {1 // (s + Csky / Ms)} · {(Cs / Ms) s + (Ks / Ms)} is described as G, and modal parameters (CskyB, CskyR, CskyP, CsB, CsR, CsP, Assuming that KsB, KsR, KsP) are GB, GR, GP, each bounce rate is dB, roll rate is dR, and pitch rate is dP, dB to dP can be calculated as the following values.
dB = GB · dxsB
dR = GR · dxsR
dP = GP · dxsP
From the above, the state estimation on the spring in the actual vehicle can be achieved based on the stroke speed of each wheel.

(Spring control unit)
Next, the configuration of the sprung mass damping control unit 33 will be described. As shown in FIG. 2, the sprung mass damping control unit 33 includes a skyhook control unit 33a that performs posture control based on the above-described sprung speed estimation value, and a frequency response that suppresses sprung vibration based on the road surface input frequency. And a control unit 33b.

[Configuration of Skyhook Control Unit]
The vehicle control apparatus according to the first embodiment includes the engine 1, the brake 20, and the S / A 3 as actuators for achieving sprung posture control. Of these, the skyhook control unit 33a controls bounce rate, roll rate, and pitch rate for S / A3, controls bounce rate and pitch rate for engine 1, and controls pitch for brake 20. The rate is controlled. Here, in order to assign a control amount to a plurality of actuators having different actions and control the sprung state, it is necessary to use a common control amount for each. In the first embodiment, the control amount for each actuator can be determined by using the sprung speed estimated by the traveling state estimation unit 32 described above.

The amount of skyhook control in the bounce direction is
FB = CskyB · dB
The amount of skyhook control in the roll direction is
FR = CskyR · dR
The amount of skyhook control in the pitch direction is
FP = CskyP · dP
It becomes. FB is transmitted to the engine 1 and S / A 3 as a bounce attitude control amount, and FR is a control executed only at S / A 3, and thus is transmitted to the damping force control unit 35 as a roll attitude control amount.

Next, the skyhook control amount FP in the pitch direction will be described. The pitch control is performed by the engine 1, the brake 20 and the S / A3.
FIG. 13 is a control block diagram illustrating actuator control amount calculation processing when performing pitch control according to the first embodiment. The skyhook control unit 33a is achieved by the engine 1 and the first target attitude control amount calculation unit 331 that calculates a target pitch rate that is a first target attitude control amount that is a control amount that can be used in common for all actuators. An engine attitude control amount calculation unit 332 for calculating an engine attitude control amount to be performed, a brake attitude control amount calculation unit 334 for calculating a brake attitude control amount achieved by the brake 20, and an S / A attitude control amount achieved by S / A3 And an S / A attitude control amount calculation unit 336.

  In the skyhook control of this system, since the first priority is to operate so as to suppress the pitch rate, the first target attitude control amount calculation unit 331 outputs the pitch rate as it is (hereinafter, this pitch rate is referred to as the pitch rate). It is described as a first target attitude control amount.) The engine attitude control amount calculation unit 332 calculates an engine attitude control amount that is a control amount that can be achieved by the engine 1 based on the input first target attitude control amount.

  In the engine attitude control amount calculation unit 332, a limit value for limiting the engine torque control amount according to the engine attitude control amount is set in order not to give the driver a sense of incongruity. Thus, the engine torque control amount is limited to be within a predetermined longitudinal acceleration range when converted to longitudinal acceleration. Therefore, if the engine torque control amount is calculated based on the first target attitude control amount and a value equal to or greater than the limit value is calculated, the pitch rate skyhook control amount that can be achieved by the limit value (suppressed by the engine 1). A value obtained by multiplying the pitch rate by CskyP (hereinafter referred to as an engine attitude control amount) is output. At this time, the value converted into the pitch rate in the conversion unit 332a is output to the second target attitude control amount calculation unit 333 described later. Further, the engine control unit 1 a calculates an engine torque control amount based on the engine attitude control amount corresponding to the limit value, and outputs the engine torque control amount to the engine 1.

  The second target attitude control amount calculation unit 333 calculates a second target attitude control amount that is a deviation between the first target attitude control amount and the value obtained by converting the engine attitude control amount into the pitch rate in the conversion unit 332a, and the brake attitude. It is output to the control amount calculation unit 334. In the brake attitude control amount calculation unit 334, a limit value for limiting the braking torque control amount is set in order to prevent the driver from feeling uncomfortable as in the case of the engine 1 (details of the limit value will be described later). .)

  Thereby, when the braking torque control amount is converted into the longitudinal acceleration, the braking torque control amount is limited to be within a predetermined longitudinal acceleration range (a limit value obtained from the occupant's uncomfortable feeling, the life of the actuator, etc.). Therefore, when the brake posture control amount is calculated based on the second target posture control amount and a value equal to or greater than the limit value is calculated, a pitch rate suppression amount (hereinafter referred to as a brake posture control amount) that can be achieved by the limit value. Output). At this time, a value converted into a pitch rate by the conversion unit 3344 is output to a third target attitude control amount calculation unit 335 described later. Further, the brake control unit 2 a calculates a braking torque control amount (or deceleration) based on the brake attitude control amount corresponding to the limit value, and outputs it to the brake control unit 2.

  In the third target attitude control amount calculation unit 335, a third target attitude control amount that is a deviation between the second target attitude control amount and the brake attitude control amount is calculated and output to the S / A attitude control amount calculation unit 336. . The S / A attitude control amount calculation unit 336 outputs a pitch attitude control amount corresponding to the third target attitude control amount.

  The damping force control unit 35 calculates a damping force control amount based on a bounce posture control amount, a roll posture control amount, and a pitch posture control amount (hereinafter collectively referred to as an S / A posture control amount). , S / A3.

[Brake pitch control]
Here, the brake pitch control will be described. In general, it can be said that it is preferable to perform both of the brakes 20 because both bounce and pitch can be controlled. However, the bounce control by the brake 20 generates braking force at the same time for the four wheels. Therefore, despite the low control priority, the deceleration effect is strong for the difficulty of obtaining the control effect, and the driver tends to feel uncomfortable. there were. Therefore, the brake 20 has a configuration specialized for pitch control. FIG. 14 is a control block diagram illustrating the brake pitch control according to the first embodiment. The vehicle body mass is m, the front wheel braking force is BFf, the rear wheel braking force is BFr, the height between the vehicle center of gravity and the road surface is Hcg, the vehicle acceleration is a, the pitch moment is Mp, and the pitch rate is Vp. Then, the following relational expression is established.

BFf + BFr = m · a
m · a · Hcg = Mp
Mp = (BFf + BFr) · Hcg
Here, when the pitch rate Vp is positive, that is, when the braking force is applied when the front wheel side is depressed, the front wheel side is further depressed and the pitch motion is promoted. In this case, the braking force is not applied. On the other hand, when the pitch rate Vp is negative, that is, when the front wheel side is lifted, the braking pitch moment gives a braking force to suppress the front wheel side lift. This contributes to improving the sense of security and flatness by ensuring the driver's field of view and making it easier to see the front. From the above
When Vp> 0 (front wheel sinks) Mp = 0
When Vp ≦ 0 (front wheel lift) Mp = CskyP · Vp
The amount of control is given. Accordingly, since the braking torque is generated only when the vehicle body is lifted up on the front side, the generated deceleration can be reduced as compared with the case where the braking torque is generated in both the lifting and sinking. Moreover, since the actuator operation frequency is only half, a low-cost actuator can be employed.

  Based on the above relationship, the brake attitude control amount calculation unit 334 is composed of the following control blocks. The dead zone processing code determination unit 3341 determines the sign of the input pitch rate Vp, and when it is positive, it outputs 0 to the deceleration reduction processing unit 3342 because control is unnecessary, and when it is negative, it determines that control is possible. The pitch rate signal is output to the deceleration reduction processing unit 3342.

[Deceleration feeling reduction processing]
Next, the deceleration feeling reduction process will be described. This process is a process corresponding to the limit by the limit value performed in the brake attitude control amount calculation unit 334. The square processor 3342a squares the pitch rate signal. This inverts the sign and smoothes the rise of the control force. The pitch rate square decay moment calculation unit 3342b calculates the pitch moment Mp by multiplying the squared pitch rate by the skyhook gain CskyP of the pitch term considering the square process. The target deceleration calculating unit 3342c calculates the target deceleration by dividing the pitch moment Mp by the mass m and the height Hcg between the vehicle center of gravity and the road surface.

  In the jerk threshold limiting unit 3342d, the calculated rate of change of the target deceleration, that is, whether the jerk is within a preset range of the deceleration jerk threshold and the extraction jerk threshold, and the target deceleration is the longitudinal acceleration limit value. Judgment is made whether or not it is within the range. If any threshold is exceeded, the target deceleration is corrected to a value within the jerk threshold range, and if the target deceleration exceeds the limit value, the limit is set. Set within the value. Thereby, the deceleration can be generated so as not to give the driver a sense of incongruity.

  The target pitch moment converting unit 3343 calculates the target pitch moment by multiplying the target deceleration limited by the jerk threshold limiting unit 3342d by the mass m and the height Hcg, and the brake control unit 2a and the target pitch rate converting unit 3344. Output for. A target pitch rate conversion unit 3344 divides the target pitch moment by the skyhook gain CskyP of the pitch term to convert it into a target pitch rate (corresponding to a brake posture control amount), and the third target posture control amount calculation unit 335 Output.

  As described above, for the pitch rate, the first target attitude control amount is calculated, then the engine attitude control amount is calculated, and the second target that is the deviation between the first target attitude control amount and the engine attitude control amount is calculated. A brake posture control amount is calculated from the posture control amount, and an S / A posture control amount is calculated from a third target posture control amount that is a deviation between the second posture control amount and the brake posture control amount. As a result, the pitch rate control amount performed by the S / A 3 can be reduced by the control of the engine 1 and the brake 20, so that the controllable area of the S / A 3 can be made relatively narrow, and the inexpensive S / A 3 can be reduced. Thus, sprung posture control can be achieved.

  Further, when the control amount by S / A3 is increased, the damping force basically increases. An increase in damping force means a hard suspension characteristic, so when high-frequency vibration is input from the road surface, it becomes easy to transmit high-frequency input and impairs passenger comfort (hereinafter referred to as high-frequency vibration characteristics). Described as worse.) On the other hand, it is possible to avoid the deterioration of the high-frequency vibration characteristics by suppressing the pitch rate by the actuator that does not affect the vibration transmission characteristics by the road surface input such as the engine 1 and the brake 20 and reducing the control amount of the S / A3. it can. The above effects can be obtained by determining the control amount of the engine 1 prior to S / A3 and determining the control amount of the brake 2 prior to S / A3.

[Frequency-sensitive control unit]
Next, frequency sensitive control processing in the sprung mass damping control unit will be described. In the first embodiment, the sprung speed is estimated based on the detection value of the wheel speed sensor 5 and the skyhook control is performed based on the estimated sprung speed control. However, when it is considered that the estimation accuracy cannot be sufficiently secured by the wheel speed sensor 5 or depending on the driving situation or the driver's intention, a comfortable driving state (a comfortable ride feeling softer than the vehicle body flatness) is guaranteed. Sometimes you want to. In such cases, vector control where the relationship (phase, etc.) of the sign of stroke speed and sprung speed is important, such as skyhook control, may make it difficult to achieve proper control due to a slight phase shift. Therefore, we decided to introduce frequency-sensitive control, which is sprung mass damping control according to the scalar quantity of vibration characteristics.

  FIG. 15 is a diagram in which a wheel speed frequency characteristic detected by the wheel speed sensor and a stroke frequency characteristic of a stroke sensor not mounted in the embodiment are simultaneously written. Here, the frequency characteristic is a characteristic in which the vertical axis represents the magnitude of the amplitude with respect to the frequency as a scalar quantity. Comparing the frequency component of the wheel speed sensor 5 with the frequency component of the stroke sensor, it can be understood that substantially the same scalar amount is taken from the sprung resonance frequency component to the unsprung resonance frequency component. Therefore, the damping force is set based on this frequency characteristic among the detection values of the wheel speed sensor 5. Here, the area where the sprung resonance frequency component exists is felt as if the occupant was thrown into the air by swinging the entire body of the occupant, in other words, the feeling that the gravitational acceleration acting on the occupant was reduced. The frequency region that brings about the wing region (0.5 to 3 Hz), and the region between the sprung resonance frequency component and the unsprung resonance frequency component is not a feeling that gravitational acceleration decreases, trot), the feeling that the human body jumps in small increments, in other words, the frequency region that brings up and down movement that can be followed by the whole body is the leopard region (3 to 6 Hz), and the region where the unsprung resonance frequency component exists Is not a vertical movement until the mass of the human body follows, but a bull region (6 to 6) is a frequency region in which vibration is transmitted to a part of the body such as the occupant's thigh. 23 Hz).

FIG. 16 is a control block diagram illustrating frequency sensitive control in the sprung mass damping control of the first embodiment. The band elimination filter 350 cuts noise other than the vibration component used for the main control from the wheel speed sensor value. The predetermined frequency domain dividing unit 351 divides the frequency band into a wide area, a horizontal area, and a bull area. The Hilbert transform processing unit 352 performs Hilbert transform on each divided frequency band, and converts it into a scalar quantity based on the amplitude of the frequency (specifically, an area calculated from the amplitude and the frequency band).
The vehicle vibration system weight setting unit 353 sets weights at which vibrations in the frequency bands of the fur region, the leopard region, and the bull region are actually propagated to the vehicle. The human sense weight setting unit 354 sets weights at which vibrations in the frequency bands of the fur region, the leopard region, and the bull region are propagated to the occupant.

  Here, the setting of human sense weight will be described. FIG. 17 is a correlation diagram showing human sensory characteristics with respect to frequency. As shown in FIG. 17, in the waft region, which is a low frequency region, the sensitivity of the occupant is relatively low with respect to the frequency, and the sensitivity gradually increases as the frequency shifts to the high frequency region. Note that the high frequency region above the bull region becomes difficult to be transmitted to the occupant. From the above, the human sense weight Wf of the wafe area is set to 0.17, the human sense weight Wh of the leopard area is set to 0.34 which is larger than Wf, and the human sense weight Wb of the bull area is larger than Wf and Wh. Set to 0.38. Thereby, the correlation between the scalar quantity in each frequency band and the vibration actually propagated to the occupant can be further increased. These two weighting factors may be changed as appropriate according to the vehicle concept and the passenger's preference.

The weight determining unit 355 calculates the ratio of the weight of each frequency band to the weight of each frequency band. If the weight of the wing area is a, the weight of the leopard area is b, and the weight of the bull area is c, the weight coefficient of the wing area is (a / (a + b + c)), and the weight coefficient of the leap area is (b / (a + b + c). )), And the weighting factor of the bull area is (c / (a + b + c)).
The scalar amount calculation unit 356 multiplies the scalar amount of each frequency band calculated by the Hilbert transform processing unit 352 by the weight calculated by the weight determination unit 355, and outputs a final scalar amount. The processing so far is performed on the wheel speed sensor value of each wheel.

  The maximum value selection unit 357 selects the maximum value from the final scalar amounts calculated for the four wheels. Note that 0.01 in the lower part is set to avoid the denominator becoming 0 because the sum of the maximum values is used as the denominator in the subsequent processing. The ratio calculation unit 358 calculates the ratio using the sum of the scalar value maximum values in each frequency band as the denominator and the scalar value maximum value in the frequency band corresponding to the waving region as the numerator. In other words, the mixing ratio (hereinafter simply referred to as the ratio) of the wafer region included in all vibration components is calculated. The sprung resonance filter 359 performs filter processing of about 1.2 Hz of the sprung resonance frequency with respect to the calculated ratio, and extracts a sprung resonance frequency band component representing a waft region from the calculated ratio. In other words, since the wing area exists at about 1.2 Hz, the ratio of this area is considered to change at about 1.2 Hz. Then, the finally extracted ratio is output to the damping force control unit 35, and a frequency sensitive damping force control amount corresponding to the ratio is output.

  FIG. 18 is a characteristic diagram showing the relationship between the vibration mixing ratio of the wafer region and the damping force by the frequency sensitive control according to the first embodiment. As shown in FIG. 18, the vibration level of sprung resonance is reduced by setting the damping force high when the ratio of the wing area is large. At this time, even if the damping force is set high, the ratio of the leopard area and the bull area is small, so that high frequency vibration or vibration that moves with the leopard is not transmitted to the occupant. On the other hand, when the ratio of the wing region is small, the damping force is set low, so that the vibration transmission characteristic more than the sprung resonance is reduced, the high frequency vibration is suppressed, and a smooth riding comfort is obtained.

Here, an advantage of the frequency sensitive control when the frequency sensitive control is compared with the skyhook control will be described. FIG. 19 is a diagram showing the wheel speed frequency characteristic detected by the wheel speed sensor 5 under a certain traveling condition. This is a characteristic that appears particularly when traveling on a road surface in which small unevenness such as a stone pavement continues. When Skyhook control is performed while traveling on a road surface exhibiting such characteristics, the damping force is determined by the value of the amplitude peak in Skyhook control. There is a problem that a very high damping force is set at an incorrect timing and high-frequency vibration is deteriorated.
On the other hand, when controlling based on a scalar quantity instead of a vector as in frequency sensitive control, a low damping force is set on the road surface as shown in FIG. Become. As a result, even if the amplitude of the vibration in the bull region is large, the vibration transfer characteristic is sufficiently reduced, so that deterioration of high-frequency vibration can be avoided. From the above, high-frequency vibration can be suppressed by frequency-sensitive control based on the scalar amount in a region where control is difficult due to deterioration in phase estimation accuracy even if skyhook control is performed using an expensive sensor or the like. .

(Unsprung vibration control unit)
Next, the configuration of the unsprung vibration suppression control unit will be described. As described in the conveyor vehicle of FIG. 12A, the tire also has an elastic coefficient and a damping coefficient, and therefore there is a resonance frequency band. However, since the mass of the tire is smaller than the mass on the spring and the elastic coefficient is high, it exists on the higher frequency side than the resonance on the spring. Due to this unsprung resonance component, the tire may flutter under the unsprung mass, which may deteriorate the ground contact property. In addition, fluttering under the spring may cause discomfort to the occupant. Therefore, in order to suppress the flutter due to unsprung resonance, a damping force corresponding to the unsprung resonance component is set.

  FIG. 20 is a block diagram illustrating a control configuration of unsprung vibration suppression control according to the first embodiment. The unsprung resonance component extraction unit 341 extracts a unsprung resonance component by applying a band-pass filter to the wheel speed fluctuation output from the deviation calculation unit 321b in the traveling state estimation unit 32. The unsprung resonance component is extracted from the region of approximately 10 to 20 Hz in the wheel speed frequency component. In the envelope waveform shaping unit 342, the extracted unsprung resonance component is scalarized, and the envelope waveform is shaped using the EnvelopeFilter. The gain multiplying unit 343 multiplies the scalarized unsprung resonance component by a gain, calculates an unsprung damping damping force control amount, and outputs it to the damping force control unit 35. In the first embodiment, the unsprung resonance component is extracted by applying a bandpass filter to the wheel speed fluctuation output from the deviation calculating section 321b in the running state estimating section 32. The unsprung resonance component may be extracted by applying a bandpass filter to the driving force, or the unsprung resonance component may be extracted by the running state estimation unit 32 by estimating and calculating the unsprung speed along with the sprung speed. Good.

(Configuration of damping force control unit)
Next, the configuration of the damping force control unit 35 will be described. FIG. 21 is a control block diagram illustrating a control configuration of the damping force control unit according to the first embodiment. In the equivalent viscosity damping coefficient conversion unit 35a, the driver input damping force control amount output from the driver input control unit 31, the S / A attitude control amount output from the skyhook control unit 33a, and the frequency sensitive control unit 33b output The frequency sensitive damping force control amount, the unsprung damping damping force control amount output from the unsprung damping control unit 34, and the stroke speed calculated by the running state estimation unit 32 are input, and these values are equivalent. Convert to viscous damping coefficient.

  In the damping coefficient arbitration unit 35b, the damping coefficient converted by the equivalent viscous damping coefficient conversion unit 35a (hereinafter, each damping coefficient is referred to as driver input damping coefficient k1, S / A attitude damping coefficient k2, frequency sensitive damping coefficient k3, unsprung). (Which is described as damping damping coefficient k4)), which arbitration is performed based on which damping coefficient is controlled, and a final damping coefficient is output. The control signal converter 35c converts the control signal (command current value) for S / A3 based on the attenuation coefficient and stroke speed adjusted by the attenuation coefficient adjuster 35b, and outputs the control signal to S / A3.

[Attenuation coefficient mediation section]
Next, the contents of arbitration by the attenuation coefficient arbitration unit 35b will be described. The vehicle control apparatus according to the first embodiment has four control modes. First, the standard mode assuming a state where an appropriate turning state can be obtained while driving in a general urban area, and second, a state where a stable turning state can be obtained while actively driving a winding road etc. In sport mode, thirdly, comfort mode that assumes a state of driving with priority on ride comfort, such as when starting at a low vehicle speed, and fourthly, highway mode that assumes a state of traveling at high vehicle speed on highways with many straight lines is there.

In the standard mode, priority is given to unsprung vibration suppression control by the unsprung vibration suppression control unit 34 while performing skyhook control by the skyhook control unit 33a.
In the sport mode, priority is given to driver input control by the driver input control unit 31, and skyhook control by the skyhook control unit 33a and unsprung vibration suppression control by the unsprung vibration suppression control unit 34 are performed.
In the comfort mode, the control for giving priority to the unsprung vibration damping control by the unsprung vibration damping control unit 34 is performed while performing the frequency sensitive control by the frequency sensitive control unit 33b.
In the highway mode, priority is given to driver input control by the driver input control unit 31, and control for adding the amount of unsprung vibration suppression control by the unsprung vibration control unit 34 to skyhook control by the skyhook control unit 33a is performed. To do.
Hereinafter, the adjustment of the attenuation coefficient in each mode will be described.

(Arbitration in standard mode)
FIG. 22 is a flowchart illustrating the attenuation coefficient arbitration process in the standard mode according to the first embodiment.
In step S1, it is determined whether or not the S / A attitude damping coefficient k2 is larger than the unsprung damping damping coefficient k4. If larger, the process proceeds to step S4 and k2 is set as the damping coefficient.
In step S2, a scalar amount ratio of the bull region is calculated based on the scalar amounts of the fur region, the leopard region, and the bull region described in the frequency response control unit 33b.
In step S3, it is determined whether or not the ratio of the bull area is equal to or greater than a predetermined value. If the ratio is greater than or equal to the predetermined value, there is a concern about deterioration of riding comfort due to high-frequency vibration. Set. On the other hand, if the ratio of the bull area is less than the predetermined value, even if the damping coefficient is set high, there is little fear of deterioration in riding comfort due to high-frequency vibration, so the routine proceeds to step S5 and k4 is set.

  As described above, in the standard mode, in principle, priority is given to unsprung vibration suppression control that suppresses unsprung resonance. However, when the damping force required by skyhook control is lower than the damping force required by unsprung vibration suppression control and the ratio of the bull area is large, the damping force of skyhook control is set and Avoid the deterioration of high-frequency vibration characteristics that accompanies the requirements. As a result, it is possible to obtain optimum damping characteristics according to the running state, and at the same time, it is possible to avoid a deterioration in riding comfort against high-frequency vibrations while achieving a flat feeling of the vehicle body.

(Mediation in sport mode)
FIG. 23 is a flowchart illustrating attenuation coefficient arbitration processing in the sport mode of the first embodiment.
In step S11, the four-wheel damping force distribution ratio is calculated based on the four-wheel driver input damping coefficient k1 set by the driver input control. The right front wheel driver input damping coefficient is k1fr, the left front wheel driver input damping coefficient is k1fl, the right rear wheel driver input damping coefficient is k1rr, the left rear wheel driver input damping coefficient is k1rl, and the damping force distribution ratio of each wheel is If xfr, xfl, xrr, xrl,
xfr = k1fr / (k1fr + k1fl + k1rr + k1rl)
xfl = k1fl / (k1fr + k1fl + k1rr + k1rl)
xrr = k1rr / (k1fr + k1fl + k1rr + k1rl)
xrl = k1rl / (k1fr + k1fl + k1rr + k1rl)
Is calculated by

In step S12, it is determined whether or not the damping force distribution ratio x is within a predetermined range (greater than α and smaller than β). If it is within the predetermined range, it is determined that the distribution to each wheel is substantially equal, and the process proceeds to step S13. If any one is out of the predetermined range, the process proceeds to step S16.
In step S13, it is determined whether or not the unsprung damping damping coefficient k4 is larger than the driver input damping coefficient k1. If it is determined that the unsprung damping damping coefficient k4 is larger, the process proceeds to step S15 and k4 is set as the first damping coefficient k. On the other hand, if it is determined that the unsprung damping damping coefficient k4 is equal to or less than the driver input damping coefficient k1, the process proceeds to step S14, and k1 is set as the first damping coefficient k.

In step S16, it is determined whether or not the unsprung damping damping coefficient k4 is the maximum value max that S / A3 can be set. If it is determined that the maximum value is max, the process proceeds to step S17, and otherwise, the process proceeds to step S18. move on.
In step S17, the maximum value of the four-wheel driver input damping coefficient k1 is the unsprung damping damping coefficient k4, and the damping coefficient that satisfies the damping force distribution ratio is calculated as the first damping coefficient k. In other words, a value that maximizes the damping coefficient while satisfying the damping force distribution rate is calculated.
In step S18, a damping coefficient that satisfies the damping force distribution ratio in a range where all the four-wheel driver input damping coefficients k1 are equal to or greater than k4 is calculated as the first damping coefficient k. In other words, a value that satisfies the damping force distribution ratio set by the driver input control and also satisfies the requirements of the unsprung vibration suppression control side is calculated.

  In step S19, it is determined whether or not the first attenuation coefficient k set in each of the above steps is smaller than the S / A attitude attenuation coefficient k2 set by skyhook control. Since the damping coefficient requested on the side is larger, the process proceeds to step S20 and k2 is set. On the other hand, if it is determined that k is equal to or greater than k2, the process proceeds to step S21 and k is set.

  As described above, in the sport mode, priority is given to unsprung vibration suppression control that suppresses unsprung resonance in principle. However, the damping force distribution rate required from the driver input control side is closely related to the vehicle body posture, and particularly because it is closely related to the driver's line-of-sight change due to the roll mode. The highest priority is to secure the damping force distribution ratio. In addition, with respect to the movement that causes the posture change in the vehicle body posture while the damping force distribution ratio is maintained, the sky vehicle body posture can be maintained by selecting Skyhook control with select high.

(Mediation in Conford mode)
FIG. 24 is a flowchart illustrating the attenuation coefficient arbitration process in the comfort mode according to the first embodiment.
In step S30, it is determined whether or not the frequency sensitive damping coefficient k3 is larger than the unsprung damping damping coefficient k4. If it is determined that the frequency sensitive damping coefficient k3 is larger, the process proceeds to step S32 and the frequency sensitive damping coefficient k3 is set. On the other hand, if it is determined that the frequency sensitive damping coefficient k3 is equal to or less than the unsprung damping damping coefficient k4, the process proceeds to step S32 to set the unsprung damping damping coefficient k4.

  As described above, in the comfort mode, priority is given to unsprung resonance control that basically suppresses unsprung resonance. Originally frequency sensitive control was performed as sprung mass damping control, and the optimum damping coefficient was set according to the road surface condition, so it was possible to achieve control that ensured riding comfort and lack of grounding feeling due to fluttering under the spring. Can be avoided by unsprung vibration suppression control. In the comfort mode, as in the standard mode, the attenuation coefficient may be switched according to the bull ratio of the frequency scalar quantity. As a result, the ride comfort can be further ensured in the super comfort mode.

(Arbitration in highway mode)
FIG. 25 is a flowchart illustrating the attenuation coefficient arbitration process in the highway mode according to the first embodiment. Since steps S11 to S18 are the same as the arbitration process in the sport mode, the description thereof is omitted.
In step S40, the S / A attitude attenuation coefficient k2 by the skyhook control is added to the first attenuation coefficient k that has been adjusted up to step S18, and is output.

  As described above, in the highway mode, the attenuation coefficient is adjusted using a value obtained by adding the S / A attitude attenuation coefficient k2 to the adjusted first attenuation coefficient k. Here, the operation will be described with reference to the drawings. FIG. 26 is a time chart showing a change in attenuation coefficient when traveling on a wavy road surface and an uneven road surface. For example, when trying to suppress the movement that the vehicle body moves swaying due to slight road surface undulations when driving at high vehicle speeds, it is necessary to detect slight wheel speed fluctuations if trying to achieve only with Skyhook control Therefore, it is necessary to set the skyhook control gain to be quite high. In this case, the movement that fluctuates can be suppressed, but if the road surface is uneven, the control gain is too large and excessive damping force control may be performed. As a result, there is a concern about deterioration in ride comfort and vehicle body posture.

  On the other hand, since the first damping coefficient k is always set as in the highway mode, a certain amount of damping force is always secured, and the vehicle body fluctuates even when the damping coefficient by the skyhook control is small. Such movement can be suppressed. Further, since it is not necessary to increase the skyhook control gain, it is possible to appropriately deal with road surface irregularities by using a normal control gain. In addition, since the skyhook control is performed with the first damping coefficient k set, unlike the damping coefficient limit, the damping coefficient decreasing process can be performed in the semi-active control region, and at the time of high-speed traveling It is possible to ensure a stable vehicle posture.

(Mode selection process)
Next, a mode selection process for selecting each travel mode will be described. FIG. 27 is a flowchart illustrating a mode selection process based on the running state in the attenuation coefficient arbitration unit of the first embodiment.
In step S50, it is determined whether or not the vehicle is in the straight traveling state based on the value of the steering angle sensor 7. If it is determined that the vehicle is traveling straight, the process proceeds to step S51. If it is determined that the vehicle is turning, the process proceeds to step S54. move on.
In step S51, it is determined based on the value of the vehicle speed sensor 8 whether or not the vehicle speed is equal to or higher than a predetermined vehicle speed VSP1 representing a high vehicle speed state. If it is determined that the vehicle speed is VSP1 or higher, the process proceeds to step S52 and the standard mode is selected. On the other hand, if it is determined that it is less than VSP1, the process proceeds to step S53 and the comfort mode is selected.
In step S54, based on the value of the vehicle speed sensor 8, it is determined whether or not the vehicle speed is equal to or higher than a predetermined vehicle speed VSP1 representing a high vehicle speed state. If it is determined that the vehicle speed is VSP1 or higher, the process proceeds to step S55 to select the highway mode. On the other hand, if it is determined that it is less than VSP1, the process proceeds to step S56 to select the sport mode.

  In other words, when driving at a high vehicle speed in a straight running state, the standard mode is selected to stabilize the vehicle body posture by skyhook control and to suppress the high frequency vibration such as leopard and bull. In addition, unsprung resonance can be suppressed. Further, when traveling at a low vehicle speed, by selecting the comfort mode, it is possible to suppress unsprung resonance while suppressing the input of vibrations such as leopard and bull to the occupant as much as possible.

  On the other hand, when the vehicle is traveling at a high vehicle speed in a turning traveling state, by selecting the highway mode, control is performed by a value obtained by adding a damping coefficient, so that basically a high damping force can be obtained. As a result, even when the vehicle speed is high, unsprung resonance can be suppressed while positively securing the vehicle body posture during turning by driver input control. In addition, when driving at low vehicle speeds, the sport mode is selected, so that the vehicle posture during turning is positively secured by driver input control, and unsprung resonance is suppressed while skyhook control is performed as appropriate. Can travel in a stable vehicle posture.

  As for the mode selection process, the control example in which the driving state is detected and automatically switched is shown in the first embodiment. However, for example, a changeover switch that can be operated by the driver is provided to select the driving mode. You may control to. As a result, ride comfort and turning performance according to the driving intention of the driver can be obtained.

As described above, Example 1 has the following effects.
(1) A damping force control unit 35 (actuator control means) that controls S / A3 (actuator) capable of suppressing the pitch motion of the vehicle, and a front and rear wheel speed difference calculation unit 312a that detects a wheel speed difference Δω between the front and rear wheels ( Wheel speed difference detecting means), and the detected fluctuation dω of the detected wheel speed difference Δω is calculated. Based on the calculated fluctuation dω, the driver input damping force control amount F (pitch rate suppression control amount) is calculated, and the damping force A front / rear slip ratio difference variation calculation unit 312c and a driver input damping force control amount calculation unit 312e (pitch rate suppression control unit) output to the control unit 35 (actuator control unit) are provided.
Therefore, pitch motion can be controlled with an inexpensive configuration without using a plurality of acceleration sensors.

(2) A damping force control unit 35 (actuator control means) that controls S / A3 (actuator) that can suppress the pitch motion of the vehicle, and front and rear wheels that calculate a difference ΔS between the slip ratio (slip amount) S of the front and rear wheels. A speed difference calculation unit 312a (slip amount difference calculation means) and a variation dS of the detected slip amount difference ΔS are calculated, and a driver input damping force control amount F (pitch rate suppression control amount) is calculated based on the calculated variation dS. And a front / rear slip ratio difference variation calculation unit 312c and a driver input damping force control amount calculation unit 312e (pitch rate suppression control unit), which output to the damping force control unit 35 (actuator control unit). .
Therefore, pitch motion can be controlled with an inexpensive configuration without using a plurality of acceleration sensors.

(3) The vehicle is an FR vehicle (rear wheel drive vehicle), and is provided with a braking / driving determination unit 312h (acceleration / deceleration determination means) that determines the acceleration direction and the deceleration direction, and a front / rear slip ratio difference variation calculation unit 312c and driver The input damping force control amount calculation unit 312e (pitch rate suppression control unit) determines the driver input damping force control amount F (pitch rate suppression control amount) based on the determination result of the braking / driving determination unit 312h (acceleration / deceleration determination unit). To do.
Therefore, pitch motion can be controlled without using a plurality of acceleration sensors even in a rear-wheel drive vehicle in which the sign of fluctuations in the front-rear wheel speed difference (slip rate difference) is the same during acceleration and deceleration. .

(4) The driver input damping force control amount calculation unit 312e (pitch rate suppression control means) multiplies the calculated variation dS by the first and second gains (predetermined gains) to calculate the driver input damping force control amount F. (Pitch rate suppression control amount) is calculated.
As described above, the driver input damping force control amount F (pitch rate suppression control amount) can be calculated by simply multiplying the fluctuation amount dS by the gain, so that the calculation load is small and it is easy to cope with the specification change.

(5) The front and rear wheel speed difference calculation unit 312a (wheel speed difference detection means) is configured to calculate the front and rear wheel speeds based on the added values of the left and right front wheel speeds ωfr and ωfl and the left and right rear wheel speeds ωrr and ωrl. The speed difference Δω is detected.
Therefore, by removing the wheel speed difference caused by the roll motion, it is possible to improve the detection accuracy of the variation dS of the front and rear wheel speed difference (slip rate difference) correlated with the pitch motion and to improve the control accuracy of the pitch motion. it can. It is also possible to configure the filter to remove the wheel speed difference caused by the pitch movements. This is because the component of the force acting between the road surface and the tire can be detected with high accuracy by removing the wheel speed fluctuation accompanying the suspension stroke.

(6) The driver input damping force control amount calculation unit 312e (pitch rate suppression control means) has a damping force control amount corresponding to the threshold value (predetermined pitch rate suppression control amount) regardless of the fluctuation dS when the vehicle speed is equal to or lower than the predetermined vehicle speed VSP1. Is calculated.
Therefore, the pitch rate can be suppressed by applying the damping force even when the fluctuation dS is small in the state from the deceleration to the vehicle stop, or in the initial drive after starting from the vehicle stop.

1 Engine 1a Engine controller (engine control unit)
2 Brake control unit 2a Brake controller (brake control unit)
3 S / A (Damping force variable shock absorber)
3a S / A controller 5 Wheel speed sensor 6 Integrated sensor 7 Rudder angle sensor 8 Vehicle speed sensor 20 Brake 31 Driver input control unit 312a Front and rear wheel speed difference calculation unit (wheel speed difference detection means)
312c Front / rear slip ratio difference fluctuation calculation unit (pitch rate suppression control means)
312e Driver input damping force control amount calculation unit (pitch rate suppression control means)
32 Traveling state estimation unit 33 Sprung vibration suppression control unit 33a Skyhook control unit 33b Frequency sensitive control unit 34 Unsprung vibration suppression control unit 35 Damping force control unit (actuator control means)

Claims (6)

Actuator control means for controlling an actuator capable of suppressing the pitch motion of the vehicle;
A wheel speed difference detecting means for detecting a wheel speed difference between the front and rear wheels;
Calculating the fluctuation of the detected wheel speed difference, calculating a pitch rate suppression control amount based on the calculated fluctuation, and providing a pitch rate suppression control means for outputting to the actuator control means,
The pitch rate suppression control means calculates a predetermined pitch rate suppression control amount regardless of the fluctuation when the vehicle speed is equal to or lower than a predetermined vehicle speed. Actuator control means for controlling an actuator capable of suppressing the pitch motion of the vehicle;
Slip amount difference calculating means for calculating the difference between the slip amounts of the front and rear wheels;
Calculating the fluctuation of the calculated slip amount difference, calculating a pitch rate suppression control amount based on the calculated fluctuation, and providing a pitch rate suppression control means for outputting to the actuator control means,
The pitch rate suppression control means calculates a predetermined pitch rate suppression control amount regardless of the fluctuation when the vehicle speed is equal to or lower than a predetermined vehicle speed. Actuator control means for controlling an actuator capable of suppressing the pitch motion of the vehicle;
A wheel speed difference detecting means for detecting a wheel speed difference between the front and rear wheels;
Calculating the fluctuation of the detected wheel speed difference, calculating a pitch rate suppression control amount based on the calculated fluctuation, and providing a pitch rate suppression control means for outputting to the actuator control means,
The vehicle is a rear wheel drive vehicle;
Acceleration / deceleration determination means for determining an acceleration direction and a deceleration direction of the vehicle based on a brake operation or an accelerator operation is further provided,
The pitch rate suppression control means, wherein in response to the determination result deceleration determination unit determines the sign of the change, computing the pitch rate suppression control amount based on the change which the code is determined A vehicle control device. Actuator control means for controlling an actuator capable of suppressing the pitch motion of the vehicle;
Slip amount difference calculating means for calculating the difference between the slip amounts of the front and rear wheels;
Calculating the fluctuation of the calculated slip amount difference, calculating a pitch rate suppression control amount based on the calculated fluctuation, and providing a pitch rate suppression control means for outputting to the actuator control means,
The vehicle is a rear wheel drive vehicle;
Acceleration / deceleration determination means for determining an acceleration direction and a deceleration direction of the vehicle based on a brake operation or an accelerator operation is further provided,
The pitch rate suppression control means, wherein in response to the determination result deceleration determination unit determines the sign of the change, computing the pitch rate suppression control amount based on the change which the code is determined A vehicle control device. The vehicle control device according to claim 1 or 3,
The wheel speed difference detecting means detects a wheel speed difference between front and rear wheels based on an added value of wheel speeds of left and right front wheels and an added value of wheel speeds of left and right rear wheels. The vehicle control device according to any one of claims 1 to 5,
The pitch rate suppression control means calculates the pitch rate suppression control amount by multiplying the calculated fluctuation by a predetermined gain.

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