JPH0732837A - Suspension system for vehicle - Google Patents

Abstract

PURPOSE:To provide a suspension system for vehicle capable of high- dimensionally ensuring the compatibility of the inexpensiveness with the controllability of the suspension characteristics. CONSTITUTION:The suspension system comprises a damper on the wheel of which the damping force is variable. According to the output signals from a first vertical acceleration sensor 6 provided at a substantially central position of a substantially frontward position of the vehicle, a second vertical acceleration sensor 7 provided at a substantially central position of a substantially rearward position of the vehicle, and a rudder angle sensor 3 for detecting the rudder angle of the vehicle, the suspension system operates as follows. Namely, it performs the control of preventing the bounce and pitch movements of the vehicle by feedback control, according to the output signals from the two vertical acceleration sensors, and performs the vehicle-turning control by feedforward control according to the rudder angle signals.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle suspension device provided with a variable damper whose damping force can be changed by an actuator, and more particularly, to reduce the number of sensors for detecting vertical acceleration. The present invention relates to a suspension device whose cost is measured.

[0002]

2. Description of the Related Art Conventionally, as a suspension device for a vehicle, as disclosed in, for example, Japanese Patent Application Laid-Open No. 3-182826, a fluid cylinder is provided between a vehicle body and each wheel. A so-called full active control suspension device (AC) in which the flow rate to the vehicle is controlled by a flow control valve independently for each wheel to make the suspension characteristics of the vehicle variable according to the operating state.
S device) is known.

However, this full active control suspension device has the drawbacks that the system becomes large in scale and the whole system becomes extremely expensive due to the use of high pressure hydraulic pressure. Therefore, a so-called "semi-active suspension device" equipped with a damping force variable damper capable of changing the damping force.
Is proposed. This variable damper uses a fluid cylinder whose interior is separated into two chambers, communicates the two chambers with an orifice, and controls the throttle amount by the orifice. The amount of restriction of the orifice is controlled as follows. That is, for example, a plurality of holes are provided in two discs,
The two discs are overlapped with their centers aligned and positioned at the boundary between the two chambers in the cylinder. Then, one of the discs is fixed, the other disc is rotated by a step motor or the like, and the orifices are formed by the holes provided in the one disc and the holes provided in the other disc being overlapped with each other. It will be. The holes provided in the one disk are made into a plurality of sizes, and the holes provided in the other disk are also made into a plurality of sizes. Therefore, the rotation position of the step motor represents the diaphragm amount, that is, the damping force.

[0004]

Since the full active suspension control device uses high pressure hydraulic pressure, the vehicle height can be positively raised and lowered, and accurate posture control can be realized.
On the other hand, the "semi-active suspension controller"
Since only the damping force is changed, the accuracy of vehicle height control (posture control) is lower than that of the full active suspension control device, but it is advantageous in cost.

However, the conventional suspension device inevitably requires three vertical acceleration sensors in order to perform the bounce control, the pitch control and the roll control. This has led to higher cost of the suspension device. The present invention is to provide a vehicle suspension device that secures both low cost and suspension characteristic control at a high level.

[0006]

In order to achieve the above object, the present invention relates to a vehicle suspension device having a wheel having a damper having a variable damping force, which is provided substantially at the center of the front portion of the vehicle. 1, an up-down acceleration sensor, a second up-down acceleration sensor provided substantially at the center of the rear portion of the vehicle, a steering angle detecting means for detecting a steering angle of the vehicle, and the first and second up-down acceleration sensors. Feedback control is performed to prevent bounce and pitch movements of the vehicle based on the output signal of the directional acceleration sensor, and control means for controlling the turning of the vehicle by feedforward control is received in response to the steering angle signal of the detection means. It is characterized by including.

[0007]

According to the active suspension device having the above structure, the number of acceleration sensors can be reduced to two. While the number of acceleration sensors can be reduced, the sensors provided are lined up in the front-rear direction of the vehicle body, so that bounce control and pitch control can be sufficiently performed with signals from these two acceleration sensors. As for the roll motion, the turning control by the feedforward control is substituted by the signal from the steering angle sensor.

[0008]

EXAMPLES Hereinafter, three preferred examples (first to third examples) to which the present invention is applied will be described. <Outline> In the suspension device of the above embodiment, the damper having the characteristic shown in FIG. 2 and the damper having the characteristic shown in FIG. 3 are used. In the damper characteristic of FIG. 2, the damping characteristic can be changed independently in both the extension direction and the contraction direction. Therefore, this damper has a small number of stages (27 stages in the embodiment).
Since the damping characteristic can be changed over the entire range, it can be applied to damping characteristic control according to the so-called “skyhook model”. Hereinafter, such a damper will be referred to as an "SD" (skyhook damper) damper. Further, in the damper characteristic of FIG. 3, when the damping characteristic in the extending direction is controlled, the characteristic in the contracting direction also changes, so that only a rough step number (five steps in the embodiment) can be set.
Therefore, it cannot be applied to the damping characteristic control according to the “Skyhook model”, and is hereinafter referred to as “AD” (adaptive damper) damper for convenience. The SD damper needs to use a high-speed step motor (actuator) because it is necessary to switch the attenuation characteristics of 27 steps in high speed. On the other hand, since the AD damper has only five stages, a low speed type step motor (actuator) is sufficient.

FIG. 1 shows what kind of signal is commonly input in order to execute suspension characteristic control in the above-mentioned three embodiments. That is, the control means of these embodiments has a vehicle speed signal VB and a steering wheel steering angle signal θH.
The vertical acceleration signal G and the brake signal Br are input. The control means inputs these signals and outputs S
Either "skyhook control" for controlling the D damper or "damping force switching control" for controlling the AD damper is performed. In the first embodiment, as shown in FIGS. 7 and 8, SD dampers are used for the left and right front wheels, and AD dampers are used for the left and right rear wheels. Also, in the second embodiment, as shown in FIGS. 16 and 17, AD dampers are used for the left and right front wheels, and SD is used for the left and right rear wheels.
It uses a damper. In the first and second embodiments, the high speed S
High-speed feedback control using the skyhook model is used for the D damper, and vehicle speed VB, for the low-speed AD damper.
Since feed-forward control is used with the steering wheel angle θH as a parameter, it is possible to secure high-performance suspension performance at a low cost. Especially the first
In the embodiment, the front suspension aims to stabilize the posture (5db in the low frequency region) and ensure the riding comfort, and the rear suspension also aims to secure the riding comfort. In the third embodiment, SD dampers are used for the left and right front wheels and the left and right rear wheels, as shown in FIGS. In this third embodiment, pitch movement, bounce movement,
Among the roll motions, feedback control using the skyhook model is applied to bounce control and pitch control, but feedforward control using the steering angle and vehicle speed is applied to roll motion control (steering control). Therefore, the acceleration sensor is not required for roll motion control, which contributes to cost reduction. : In the first to third embodiments, the controls "G through control" and "large amplitude input control" are commonly performed. -1: "G through control" (FIG. 33) corrects the damper characteristics in the soft direction to ensure a comfortable ride when there is a vertical G motion with a relatively large amplitude. If the motion matches during high-speed driving or turning, the correction in the soft direction is limited to ensure steering stability. -2: Also, in the "large amplitude input control" (FIG. 29), when there is an up and down G motion of even larger amplitude, the damping force is increased with the first consideration being to ensure steering stability. When such a vertical G movement matches during high-speed driving or turning, the correction in the hardware direction is further enhanced to ensure further steering stability. -3: What is "G through control" and "large amplitude input control"?
Since the vertical G motion is a problem, they may interfere with each other, but in this embodiment, there is no problem because the "large amplitude input control" is given a higher priority than the "G through control". <Damper> Figures 2 and 3 show SD damper characteristics and AD, respectively.
Shows damper characteristics. The horizontal axis represents the rotational position P of the step motor, and the vertical axis represents the damping force. Further, as described above, the SD damper is set to 27 positions and the AD damper is set to 5 positions.

In the SD damper of FIG. 2, when the rotational position P moves in the positive direction, the damping force in the extension direction increases, but the damping force in the contraction direction becomes small. Further, when the rotational position P is at the negative position, the damping force in the extension direction becomes small, but the damping force in the contraction direction increases. That is, if the stepper motor is in a more positive position, the movement to raise the vehicle body is
The damping force in the extension direction acts to have a so-called "hard" characteristic, but to a movement for lowering the vehicle body, a so-called "soft" characteristic. Further, if the step motor is in a more negative position, a damping force in the direction of contraction acts on the movement to lower the vehicle body, which is a so-called "hard" characteristic, but tries to raise the vehicle body. It has a so-called "soft" characteristic for exercise.

On the other hand, the AD damper, as shown in FIG.
The greater the rotational position P of the step motor, the more
It has a "hard" characteristic in both the extension direction and the contraction direction, and a "soft" characteristic as the rotational position becomes smaller. <Overall Configuration of Control System> In common with the first to third embodiments, "steering control" relating to steering stability, and making suspension characteristics hard when a vertical acceleration having a large amplitude is detected. "Large input amplitude control" to improve safety by using "G-through" to make such acceleration "through" by softening the suspension characteristics when high-frequency vertical acceleration (acceleration when driving on a rough road) is detected. "Control", "small amplitude control" that makes suspension characteristics relatively soft when a vertical acceleration having a small amplitude is detected, and the like are implemented. 4 and 5 are maps that schematically show the relationship between the priority order of these controls and the operating regions. In FIG. 5, the horizontal axis represents the vertical acceleration G of the vehicle body, and the vertical axis represents the vehicle speed VB.

FIG. 6 is a control block diagram showing the mutual relationship of the above various controls. <First Embodiment> FIG. 7 shows a first embodiment in which an SD damper (not shown) is used for the front wheel suspension and an AD damper (not shown) is used for the rear wheel suspension. In the figure, 2L and 2R are vertical acceleration sensors provided on the left and right, respectively, and generate acceleration signals GL and GR. This acceleration signal is sent to the controller 10. Further, the left and right SD dampers are respectively driven by the high speed step motors 1FL and 1FR, and the rear wheel left and right AD dampers are respectively driven by the low speed step motors 1R.
L and 1RR drive. The steering angle θH is detected by the steering angle sensor 3 and sent to the controller 10. Also, the vehicle speed sensor 4
The vehicle speed VB detected by and the signal BR detected by the brake switch 5 are sent to the controller 10, respectively.

FIG. 8 is a block diagram showing the outline of the control of the first embodiment. As shown in FIG. 8, the steering control (steering control) SH using the skyhook model has acceleration signals GL and GR, a steering angle signal θH, a brake signal BR, and a vehicle speed signal V.
Input B etc. to control the left and right SD dampers of the front wheels. The actual operation of the control unit SH is realized by the controller 10 executing the control procedure according to the flowchart of FIG. Further, the two damping force switching control units AA are
The steering angle signal θH, the brake signal BR, and the vehicle speed signal VB are input to control the two AD dampers for the left and right rear wheels. Control unit A
The actual operation of A is realized by the controller 10 executing the control procedure according to the flowchart of FIG. As described above, since the feedback control that requires high-speed operation is not applied to the control of the rear wheels, the expensive upper and lower G sensors are not required on the rear wheel side, and the inexpensive low-speed AD damper is not required. Will be enough.

Further, both the SD damper for the front wheels and the AD damper for the rear wheels have "G through control" (FIG. 33) and "large input amplitude control" (FIG. 29) which are performed based on the up and down G signals. Applied. Front wheel suspension control (SH / stability control): 1st implementation
The control procedure of the example control unit SH is shown in FIG. The safety control of the first embodiment will be described with reference to the flowchart of FIG.

In step S2, the acceleration signals GL,
Various signals such as GR, steering angle signal θH, brake signal BR, vehicle speed signal VB, etc. are input. In step S4, the acceleration signal G
Integrate L and GR respectively to obtain the vertical velocity VGL and VG of the vehicle body.
Find R. In step S6, based on the vehicle speed VB, as shown in FIG.
A speed threshold value VG0 is obtained according to the characteristic diagram of 0. In step S8, the vertical speed VGn (n indicates L left and R right)
Are respectively compared with the threshold value VG0. If | VGn | ≧ VG0 is determined in step S8, the process proceeds to step S12. On the contrary, if | VGn | <VG0 is determined, VGn = 0 is set in step S10, and the vehicle body is moved in the vertical direction. If not moving, then step S
Proceed to 12. As will be described later, since VGn is an important factor that determines the position of the damper, the dead zone of control is in the region of the vehicle body vertical velocity VGn where | VGn | <VG0.

In step S12, a control gain K1 for determining the damper position P is calculated. The control gain K1 shows a larger value as the vehicle speed VB is higher and the steering wheel angular velocity (= time change) θ'H (= dθH / dt) is larger. That is, when the vehicle speed VB and the steering angular speed θ'H are high, the SD damper position is changed at high speed. In step S14, the target vehicle up-down velocity VGTRn (n is L
Left and R are shown), and VGTRn = VGnK1 (1) is calculated. If this VGTRn is positive, step S20.
To the target position PFTRn (n is L
Left, R right) are calculated according to PFTRn = PFn-1 (2). The expression (2) changes the damping characteristic by one step to be hard so as to suppress the displacement of the vehicle body in the extension direction. On the contrary, when VGTRn is negative, step S
In step 18, the target position PFTRn of the front wheel SD damper is calculated according to PFTRn = PFn + 1 (3). The expression (3) changes the damping characteristic by one step so as to suppress the displacement of the vehicle body in the contraction direction. PFn in steps S18 and S20 is the current position of the step motor of the front wheel SD damper obtained in step S2.

In step S22, a limit value PLMT as shown in FIG. 12 is obtained based on the vehicle speed VB. Limit value PLMT
Has a tendency to increase as the vehicle speed VB increases, as shown in FIG. This is because it is necessary to increase the responsiveness by increasing the tolerance of control changes as the vehicle speed increases. In step S24, this limit value PLMT is compared with the target position PFTRn, and this limit value is compared with the target value PFTRn.
If is exceeded, the target value is clipped to this limit value in step S26.

In step S28, it is checked whether the flag INHIBIT is set. If this flag is not set, in step S30, the motors 1FL and 1FR are rotated so that the front wheel SD damper can achieve the target damping force. Here, the flag INHIBIT is "G through control" (step S276 in FIG. 33) or "large amplitude input control" (FIG. 2).
This is a flag that is set in step S218) of 9 and is set when there is a possibility that the front wheel suspension characteristic and the rear wheel suspension characteristic may become excessively different. Therefore, if this flag is set, step S30 is not executed and the motor position of the damper (of the front wheels) is not changed. Rear Wheel Suspension Control (First Embodiment) The suspension characteristics of the rear wheels are determined by a 5-stage AA damper. In other words, the damping force of this damper is determined by the rotational positions of the motors 1RL and 1RR. FIG. 13 shows a control procedure of the rear wheel AA damper motors 1RL and 1RR.

The rear wheel suspension control of the first embodiment is
This is determined by feedforward control so that the rear wheels tend to understeer the suspension characteristics of the front wheels determined by the control procedure of FIG. 9. The reason for using feedforward control is that the control speed is fast and that the AA damper does not require advanced feedback control.

That is, in step S40, the vehicle speed VB
The target damping force PR (VB) of the rear wheels corresponding to the above is determined according to the characteristics shown in FIG. Here, in FIG. 14, the solid line shows the damping characteristics of the rear wheels while the vehicle speed VB is increasing, and the broken line shows the damping characteristics of the rear wheels while the vehicle speed is decreasing. The characteristics are such that when the vehicle speed is decelerating, a lower damping force can be obtained compared to when increasing the vehicle speed. The characteristic of FIG. 14 is to obtain a more understeer tendency in order to stabilize the posture of the vehicle body during deceleration.

In step S42, the current steering angle θH is compared with a predetermined threshold value θH0 to determine whether or not the vehicle is turning. When not turning (| θH | <θH0),
The process proceeds to step S52 and below. In step S52, it is determined whether or not the "large amplitude input control" is currently executed (F =
2) is checked, and step S54 is to check the flag (F = 3) as to whether or not the "G through control" is currently being executed. When neither the "large amplitude input control" nor the "G through control" is executed, the process returns from the step S54 to the main routine. Therefore, the step S30 of FIG.
When is executed, the target position PR set in step S40 is set to the rear wheel damper motors 1RL and 1RR.

If the vehicle is turning (| θH | ≧ θH0), the process proceeds to step S44 and thereafter. In step S44, the current motor position of the front wheel damper on the outside of the turn is monitored. This motor position is POF. In step S46, FIG.
According to the characteristics of 5, the target damping force of the rear wheel damper (that is,
Motor position PRTRn) is determined. The characteristic of FIG. 15 is such that a damping force lower than the damping force POF of the outer turning wheel obtained in step S44 is set on the rear wheel side. Since the SD damper on the front wheel side has 27 steps and the AA damper on the rear wheel side has 5 steps, the horizontal axis of FIG. 15 indicates the relative damping strength of the rear wheel damper to the front wheel damper. That is, for example,
When the damping force of the front wheels is a damping force corresponding to the fourth step of the damping force of the rear wheel side damper (for example, the 25th step), the damping force of the rear wheel side damper is one step lower than the fourth step. It is to reduce it to 3 steps.

The characteristics shown in FIG. 15 may be changed to take the vehicle speed into consideration. In step S48, step S40
The target damping force PR obtained in accordance with the vehicle speed VB and the step S4
In step 6, the target damping force PRTRn calculated in relation to the damping force of the front wheels is compared. If the latter is large (PR ≤ PRTR
n), in step S50, the target value of the rear wheel damping force is set as the damping force PR obtained in relation to the vehicle speed. That is, PRTRn = PR (4) This PRTRn is set in the step motor of the rear wheels at the time when step S30 of FIG. 9 is executed. On the other hand, if it is determined in step S48 that PR> PRTRn, the target value PRTRn obtained in step S46 is set in the motor in relation to the damper characteristics of the front wheels. This is done to ensure that the rear wheel suspension control of the first embodiment has an understeer characteristic during turning. That is, for example, when the damping force PR determined in step S40 according to the vehicle speed VB is "third stage" and the damping force PRTRn determined in step S46 is "second stage", PRTRn Since <P R, if P R is adopted as the damping force, the relationship that the rear wheels are understeer with respect to the front wheels may not be established. Therefore, only when PR ≦ PRTRn in step S48, that is, only when the understeer relationship is secured, the PR determined in step S50 according to the vehicle speed VB is adopted as the rear wheel damping force.

If it is determined in step S52 that "large amplitude input control" is in progress (F = 2), then in step S60
Increase the damping force by one step. During "Large amplitude input control" while traveling straight ahead, the rear wheel side also has the damper characteristic on the hard side so that it can cope with vertical acceleration with a large acceleration when riding on an obstacle rather than maintaining the understeer characteristic. It is necessary to increase In step S62, the control for correcting the characteristic of the rear wheel to the hard side is continued for a predetermined time. This is because, as will be described later, in the "large amplitude input control" (Fig. 29), the suspension characteristics of the front wheels are also changed to the hard side for a predetermined time.

When the "G through control" is being executed while the vehicle is traveling straight (YES in step S54), the step S is performed.
At 56, the rear wheel damper power is reduced by one step. The reason for lowering by one step is that the suspension characteristics of the front wheels are also changed to the soft side in the “G through control” (FIG. 33), as described later. Effects of the First Embodiment As described above, according to the suspension device of the first embodiment, the rear wheel side damper employs an AA damper having a small number of steps, and the damper is driven at a low speed step. Motor (1
Since RL, 1RR) is adopted, the cost can be reduced.
Feedback control based on an acceleration signal that requires high speed control is difficult to apply to such a low speed damper, so in this first embodiment, feed forward control based on the vehicle speed signal VB (step S40 in FIG. 13). ) Is adopted. By adopting this feedforward control, the control can be simplified, which contributes to cost reduction. The feedforward control and the adoption of the low speed damper may reduce the maneuverability, but in this first embodiment, the high speed SD damper is adopted on the front wheel side and the skyhook control based on the vertical acceleration signal is adopted. (Step S12 of FIG.
Since step S14) is carried out, it is possible to secure maneuverability. : When determining the damper characteristics of the rear wheels during turning, the damper characteristics of the front wheels are referred to. This is because the damper characteristics of the rear wheels during turning affect the steerability, and therefore the damper characteristics of the rear wheels should be determined so as to establish a predetermined relationship with the damper characteristics of the front wheels. Particularly in this embodiment, the damper characteristics of the rear wheels are determined so that the rear wheels maintain an understeer tendency with respect to the front wheels (FIG. 1).
3 step S48, step S50). : The setting of the control gain when determining the damper characteristics of the front wheels and the rear wheels is set so that the vehicle speed VB becomes higher, the vehicle speed is decelerated (characteristics of the broken line in FIG. 14), and the steering angular speed becomes larger (K1 in FIG. ), It is set to be hard. This is because the higher the vehicle speed VB, the slower the vehicle speed, and the higher the steering angular velocity, the higher the suspension characteristics and the better the responsiveness. <Second Embodiment> In the first embodiment, the SD damper is used for the front wheels and the AD damper is used for the rear wheels. In the second embodiment, as shown in FIG. 16, an AD damper is attached to the front wheels.
SD dampers are used for the rear wheels. Therefore, in the second embodiment, the low speed stepping motor 1'FL,
1'FR is provided and high speed step motor 1'R is provided on the rear wheels.
L and 1'RR are provided. FIG. 17 shows a control block diagram of this second embodiment.

FIG. 18 shows rear wheel suspension control, and FIG. 19 shows front wheel suspension control. 18 is substantially similar to FIG. 9 of the first embodiment for performing SD damper control on the rear wheels, and FIG. X is for the first embodiment for performing AD control on the front wheels in the embodiment. 13 is substantially similar to FIG. Rear Wheel Suspension Control (Second Embodiment) The steering control for the rear wheels of the second embodiment will be briefly described with reference to the flowchart of FIG.

In step S60, the vertical acceleration signals GL and GR at the rear wheel position and the steering angle signal θ
Various signals such as H, brake signal BR and vehicle speed signal VB are input. In step S62, the acceleration signals GL and GR are integrated to obtain the vertical velocity VG at the rear wheel position of the vehicle body.
Calculate L and VGR. In step S64, the speed threshold value VG0 is obtained based on the vehicle speed VB according to the characteristic diagram shown in FIG. In step S66, the vertical speed VGn (n is L
Left and R right) are respectively compared with the threshold value VG0. If | VGn | ≧ VG0 is determined in step S66, the process proceeds to step S70. On the other hand, if | VGn | <VG0 is determined, VGn = 0 is set in step S68, whereby the vehicle body moves in the vertical direction. If not, move to step S70.

In step S70, the control gain K1 for determining the damper position P is calculated. This control gain is such that the steering wheel angular velocity θ'H (= dθ increases as the vehicle speed VB increases.
The larger H / dt), the larger the value. In step S72, a target vehicle up-down velocity VGTRn for canceling the up-down movement velocity of the vehicle body at the present time is calculated as VGTRn = VGnK1 (5). If this VGTRn is positive, step S76.
Then, the target position PRTRn of the rear wheel SD damper is calculated according to PRTRn = PRn-1 (6). The expression (6) changes the damping characteristic by one step to be hard so as to suppress the displacement of the vehicle body in the extension direction. On the contrary, when VGTRn is negative, step S
Proceeding to 76, the target position PRTRn of the rear wheel SD damper is calculated according to PRTRn = PRn + 1 (7). The expression (7) changes the damping characteristic by one step so as to suppress the displacement of the vehicle body in the contraction direction. PRn in steps S76 and S78 is the current position of the step motor of the rear wheel SD damper obtained in step S60.

In step S80, a limit value PLMT as shown in FIG. 12 is obtained based on the vehicle speed VB. Step S
At 82, this limit value PLMT is compared with the target position PFTRn. If the target value PRTRn exceeds this limit value, the target value is clipped to this limit value at step S84. In step S86, it is checked whether the flag INHIBIT is set. If this flag is not set, in step S88, the motors 1'FL and 1'FR of the rear wheel damper are rotated so that the front wheel SD damper can achieve the target damping force. Here, the flag INHIBIT is a flag that is set in the "G through control" or the "large amplitude input control" as in the first embodiment, and these control procedures are the "G through control" or the "large through control". If the "amplitude input control" is executed as it is, there is a possibility that the front wheel suspension characteristics and the rear wheel suspension characteristics may become excessively different.This is to prevent the motor position of the front wheel damper from being changed. is there. Front Wheel Suspension Control (Second Embodiment) The suspension characteristics of the front wheels are determined by a 5-stage AA damper. In other words, the damping force of this damper is determined by the rotational positions of the motors 1'RL, 1'RR. FIG. 19 shows a control procedure for the motors 1'RL and 1'RR of the rear wheel AA damper.

The front wheel suspension control of the second embodiment is
With respect to the suspension characteristics of the front wheels, the suspension characteristics of the front wheels are determined by the feedforward control so that the suspension characteristics of the rear wheels determined by the control procedure of FIG. 18 tend to be understeered. That is, in step S100, the target damping force PF (VB) of the front wheels according to the vehicle speed VB is determined according to the characteristics shown in FIG. Here, in the characteristic diagram of FIG. 20, the solid line shows the damping characteristics of the front wheels while the vehicle speed VB is increasing, and the broken line shows the damping characteristics of the front wheels while the vehicle speed is decreasing. The characteristics are such that when the vehicle speed is decelerating, a lower damping force can be obtained compared to when increasing the vehicle speed. The characteristic of FIG. 20 is to obtain a more understeer tendency in order to stabilize the vehicle body posture during deceleration.

In step S102, the present steering angle θH is set to a predetermined threshold value θH0 in order to determine whether or not the vehicle is turning.
Compare with. If the vehicle is not turning (| θH | <θH0), the process proceeds to step S102 and thereafter. In step S102, a flag (F = 2) is checked as to whether or not the "large amplitude input control" is currently executed, and in step S104, whether or not the "G through control" is currently executed (F =
3) is investigated. When neither "large amplitude input control" nor "G through control" is executed, step S104
Since the process returns from the main routine to the main routine, the target position PF set in step S100 is set to the motor 1'FL, 1'F of the front wheel damper when step S88 in FIG. 18 is executed.
Set to R.

If the vehicle is turning (| θH | ≧ θH0), the process proceeds to step S104 and thereafter. In step S104, the current motor position of the damper of the wheel of the turning outer wheel among the rear wheels is monitored. This motor position is POF. Step S
At 106, the target damping force of the damper of the front wheels (that is, the motor position PFTRn) is determined according to the characteristics of FIG. Figure 2
The first characteristic is a characteristic that a damping force higher than the damping force POF of the turning outer wheel obtained in step S104 is set on the front wheel side so that the rear wheel has an understeer characteristic.

In step S108, step S100
The target damping force PF obtained according to the vehicle speed VB and the step S0
In step 6, the target damping force PFTRn calculated in relation to the rear wheel damping force is compared. If the latter is small (PF> PFTR
n), in step S110, the target value of the front wheel damping force is set to the damping force PF obtained in relation to the vehicle speed. That is, PFTRn = PF (8) This PFTRn is set in the step motor of the front wheels when step S880 of FIG. 18 is executed. On the other hand, if it is determined in step S108 that PF ≦ PFTRn, step S10 is performed in relation to the damper characteristics of the rear wheels.
The target value PFTRn obtained in 6 is set in the motor.

The suspension control of the front wheels of the second embodiment sets the suspension characteristics of the front wheels such that the damping characteristics of the rear wheels are understeer compared to the damping characteristics of the front wheels during turning. That is, for example, when the damping force PF determined according to the vehicle speed VB in step S100 is the "second stage" and the damping force PFTRn determined in step S06 is the "third stage", PFTRn Since> PF, if PF is adopted as the damping force, the relationship that the rear wheels are understeer with respect to the front wheels may not be established. Therefore, only when PR> PRTRn in step S108, that is, only when the understeer relationship is secured, the PF determined in step S110 according to the vehicle speed VB is adopted as the rear wheel damping force.

When it is determined in step S02 that "large amplitude input control" is in progress (F = 2), step S120
Then, increase the damping force by one step. "Large amplitude input control" while going straight
This is because it is necessary to increase the damper characteristic to the hard side on the rear wheel side as well, in order to cope with the vertical movement with a large acceleration when riding on an obstacle or the like, rather than maintaining the understeer characteristic. In step S122, the control for correcting the characteristics of the front wheels to the hardware side is continued for a predetermined time. This is because, as will be described later, in the "large amplitude input control" (Fig. 29), the suspension characteristics of the front wheels are also changed to the hard side for a predetermined time.

When the "G through control" is being executed while the vehicle is traveling straight (YES in step S104), the rear wheel damper force is reduced by one step in step S106. The reason for lowering by one step is that the suspension characteristics of the front wheels are also changed to the soft side in the “G through control” (FIG. 33), as described later. <Third Embodiment> In the third embodiment, SD dampers are used for the front wheels and the rear wheels, and the vertical direction G is detected by one G sensor 6 provided on the front wheel side and on the rear wheel side. This is performed by only one G sensor 7. As shown in FIG. 22, if only one sensor is used for each of the front wheels and the rear wheels, it becomes difficult to detect the roll motion of the vehicle body. However, the roll (steering) control is most necessary at the time of turning, and if necessary, for example, by making the front wheel characteristics hard, it is possible to obtain necessary and sufficient steering stability characteristics. And, in the past, three or more (one pair of sensors in the left-right direction, one sensor in front or in the back) was required, but in the third embodiment, two sensors are sufficient, so that the cost is reduced. To help.

FIG. 23 shows the overall construction of the control system of the third embodiment. Roll (steering) control is based on the steering angle signal θH
And the vehicle speed signal VB, the bounce and pitch components of the vehicle body posture are determined based on the vehicle speed signal VB and the acceleration signals from the front and rear G sensors. Bounce / Pitch Control (Third Embodiment) FIG. 24 is a control flowchart for the bounce / pitch control part performed for the SD dampers of all the wheels in the third embodiment. FIG. 25 is a flowchart showing a control procedure of roll control similarly performed for the SD dampers of all wheels.

In FIG. 24, steps S132 to S160 are substantially the same as steps S2 to S30 of FIG. 9, except that the third embodiment differs from the front G sensor in step S132. The difference is that the signal GF and the signal GR from the rear G sensor 7 are input, and in step S134, the vertical movement speed VF of the front part of the vehicle body and the vertical movement speed VR of the rear part of the vehicle body are input. A big difference from the flowchart of FIG. 24 is that when the flag F is 1 in step S130, the bounce / pitch control of steps S132 to S160 is not executed. The case where the flag F is equal to 1 becomes clear by the roll control of FIG. Roll Control FIG. 25 is a flowchart showing the control procedure of roll control (steering control) performed on the SD dampers of the front and rear wheels. In step S170, it is checked whether the vehicle body is turning (| θH | ≧ θH0) or going straight (| θH | <θH0). If the vehicle is traveling straight (NO in step S170), the temporal change θ'H (= dθ / dt) of the steering angle is checked in step S192. It is determined in step S170 that the vehicle is turning (YES), or in step S192 the steering angle has changed (N
If it is determined to be O), the process proceeds to step S172, and the flag F is set to 1 to indicate that the roll control is to be performed.
To If the vehicle is not turning and the steering angle is not changing, the flag is reset to 0 in step S194. Therefore, when the flag F is 0, the bounce / pitch control of FIG. 24 (FIG. 24) is performed for each wheel, and the roll control of FIG. 25 is not performed.
This is because, as described above, the roll (steering) control is most necessary at the time of turning in the first place, and when turning, for example, by making the front wheel characteristics hard, necessary and sufficient steering control characteristics can be obtained. Because you can.

A case where it is determined that the vehicle has started or is turning will be described. In this case, step S172 →
In step S174, the target damping force PF for the front wheels is set to the vehicle speed VB and the steering angle θH.
Based on. The target damping force PF is, for example, as shown in FIG.
It is determined based on the vehicle speed VB and the steering angle θH according to the characteristics as shown in FIG. That is, the characteristics shown in the figure are such that the higher the steering angle θH and the higher the vehicle speed VB, the greater the damping force (hardening the damper characteristics). Further, in step S178, the coefficient A is determined based on the steering angular velocity θ'H. The coefficient A has a characteristic that it becomes larger as the steering angular velocity θ′H becomes larger, as shown in FIG. 27, for example. Step S17
At 8, the target damping force PFAn (n represents right or left) is calculated.

PFAn = PF · A (9) Thus, in steps S174 to S178,
The target damping force PFAn of the front wheels is determined to have a larger value as the vehicle speed increases, the steering angle increases, and the steering angular velocity increases. In step S180, the coefficient K for the rear wheels is determined. This coefficient K is, for example, as shown in FIG. 28, a coefficient smaller than 1 and has a characteristic that it decreases as the vehicle speed VB increases.

In step S182, the target damping force PFAn for the front wheels and the actual current damping force PRn of the rear wheels are compared. The target damping force PFAn of the front wheels is the current damping force PRn of the rear wheels.
If it is greater than PFAn ≦ PRn, the process proceeds to step S184, and in order to set the target damping force PFAn to the final target damping force PFTRn of the front wheels, PFTRn = PFAN ... (10), and step In S186, PRTRn = PFAn · K (11) is set by using the coefficient K obtained in step S180 so that the rear wheels have an understeer tendency with respect to the front wheels. That is, as shown in FIG. 28, the coefficient K is a number smaller than 1. Therefore, according to the equation (11), the rear wheel is always set to be smaller than the damping force of the front wheel.

On the other hand, If the current rear wheel damping force PRn is smaller than the target front wheel damping force PFAn in step S182, the rear wheels may be oversteered, so in step S188, PFTRn = P + -... ( 12) The expression (12) means that the damping force PFTRn of the front wheels is set so that the front wheels on the outside of the turn have a hard characteristic in the contraction direction and the front wheels on the inside of the turn have a hard characteristic in the extension direction. Is to do. Further, in step S190, PRTRn = PRn · K (13) is set so that the rear wheel characteristics have a lower damping force than the current damping force so that the understeer tendency is secured. Effects of the third embodiment Thus, according to the third embodiment :: Signals from two G sensors (6, 7) provided substantially at the center in the vehicle width direction and at the front and rear in the vehicle length direction, respectively. Control for suppressing bounce / pitch based on GF, GR and signal θH from the steering angle sensor (Fig. 24)
The turning control (roll control) is performed in a feedforward manner based on the steering angle signal θH. By doing so, it is possible to reduce the number of G sensors by one as compared with the related art, and yet, it is possible to realize both the bounce / pitch control and the roll control. : The feedforward control for roll control is corrected by the vehicle speed VB and the steering angular velocity θ'H for the front wheels (step S174, step S178). : First, the damper force for the front wheels is determined, and thereafter, the damper characteristics of the rear wheels are determined so as to be understeer characteristics than the front wheels (steps S182 to S182).
190). : By using the flag F, roll control (turn control) is prioritized over bounce / pitch control.
This ensures posture control in the roll direction during turning. : In roll control, feedforward control is performed only during turning. <Large Amplitude Input Control and G-Through Control> The three embodiments (FIGS. 8, 17, and 23) have been described above. Next, the large-amplitude input control and the G-through control, which are commonly applied to the suspension devices of these embodiments, will be described. Large-amplitude input control The large-amplitude input control is performed by, for example, the vertical acceleration signal G in order to ensure safety when the vehicle body gets on an obstacle.
Is a control for detecting that a large amplitude is input and increasing the damping force.

FIG. 29 shows the control procedure of this large amplitude input control. Arbitration between the large amplitude input control (FIG. 29) and the SH control in the first embodiment, for example, is performed by the flag F described above. That is, when the large amplitude input control is performed, the flag F is set to 2 in steps S210 and S222. On the other hand, in the rear wheel control (FIG. 13) of the first embodiment, when the flag F = 2 is detected, step S60 and the subsequent steps are executed.

First, with reference to the flow chart of FIG. 29, what kind of control will be executed when the G sensor output is a large amplitude input will be described. In step S200, the signal from the G sensor is integrated to obtain the vehicle speed VG in the vertical direction. In step S202, the steering angle θH and the threshold value θH0 are compared to determine whether or not the vehicle is turning. The control is different between when it is determined that the vehicle is turning and when it is determined that the vehicle is traveling straight. Further, as will be described later, the control differs depending on the magnitude of the vehicle body speed VG.

FIG. 32 is a table summarizing the control modes of the large amplitude input control. In the figure, the control interval means, for example, a time interval in which the control procedure of FIG. 13 of the first embodiment is executed. The shorter this time interval, the quicker the control and the more sensitive the input. In the control procedure of FIG. 29, when the control interval tx is “slow”, when the control interval tx is t0>t1> t2, tx = t0 is set, and “early” is set tx = t2. , "Normal" means that tx = t1. Further, in FIG. 32, “enlarging” the upper limit value PLMT means expanding the characteristic of FIG. 12 to 1.2 times.

If the vehicle is traveling straight (| θH | <θH0), the process proceeds to step S219, and a threshold value GA for determining whether or not a large amplitude input has been made is determined. This threshold value GA is determined based on the vehicle speed VB according to the characteristic shown in FIG. 30, for example. The characteristic of the threshold GA in FIG. 30 is a threshold that increases as the vehicle speed VB increases. If the vehicle speed VG in the vertical direction is smaller than GA, that is, | VG | <GA, the process proceeds to step S230, the flag F is set to 0, and the control cycle time tx is set in step S232.
Is set to the normal interval (t1) and PLMT is not changed, so the value of "normal" (FIG. 12) is set.

When it is determined in step S220 that the vehicle body speed VG is greater than the threshold value GA, the flag F is set to 2 in step S222. Then, in step S224, the control interval is set to "early" (tx = t2),
In step S226, the upper limit value PLMT is increased 1.2 times. On the other hand, when there is a large amplitude input while the vehicle is traveling straight, it is determined that the vehicle is traveling straight in step S42 of the control procedure in FIG. 13 (rear wheel control of the first embodiment), the process proceeds to step S52, and the value of the flag F is changed. Be examined. As described above, since the flag F is set to 2 in step S222 of FIG. 29,
In step S52, YES is determined and step S60
Will proceed to. In step S60, in order to make the damping force target value PRTRn of the rear wheel damper harder than the current value, PRTRn = PRn + 1 (14). In step S62, such rear wheel damping force control is continued for a predetermined time. The reason for continuing is that such a large amplitude input state is considered to continue for the predetermined time.

In this way, when there is a large amplitude input while traveling straight ahead, the damper force control for the rear wheels (FIG. 13) and the "large amplitude input control" (FIG. 29) operate in concert. And deal with it. That is, when there is a large amplitude input while going straight,
As for the rear wheels, the damping force is increased (step S6).
0), by making the control interval between the front wheel control (FIG. 9) and the rear wheel control (FIG. 13) shorter (time interval t2), the reaction to the impact input is sharpened. Further, there are cases where the damping force must be increased by speeding up the reaction. In such a case, by increasing the upper limit value PLMT (step S226), even if the damping force becomes large due to the damping force enhancement as a response to the impact input, it is not clipped.

During turning (| θH | ≧ θH in step S202)
If it is determined to be 0), the magnitude of the impact is measured by comparing the predetermined threshold value GB with the vertical speed VG of the vehicle body in step S208. The threshold GB is determined in step S206 based on the steering wheel steering angle θH and the steering angular velocity θ′H, for example, according to the characteristic shown in FIG. This characteristic is that the larger the steering angle θH of the steering wheel and the larger the steering angular velocity θ′H, the larger the value of GB.

When the impact input applied to the vehicle body is large such that | VG | ≧ GB, the flag F is determined in step S210.
Is set to 2 and the control interval is lengthened in step S212, that is, the response of the damper control to the shock input is made slow. Then, steps S214 and S216
In the above, the target damping force is set so as not to cause a difference of three or more steps between the front wheels and the rear wheels or between the right wheels and the left wheels. When the difference between the front wheels and the rear wheels is about 3 steps or more, | Pf-Pr | ≧ 3 (15), and the difference between the right wheels and the left wheels is about 3 steps or more. Then, | PL-PR | ≧ 3 (16). However, Pf is the final target damping force PF of the front wheels
TRn, and Pr is the target damping force RTRn of the rear wheels. In such a case, the process proceeds to step S218 to output the signal INHIBIT. The signal INHIBIT is, for example, step S1 in FIG.
Reference numeral 58 is a signal for controlling whether to actually output damping force signals to the dampers for the front wheels and the rear wheels. When this signal INHIBIT is generated, the change of the damping force is stopped, so the generated damping force is generated between the front wheel and the rear wheel or the right wheel-
There will be no more than three steps between the left wheels.

On the other hand, if the impact force is small even during turning (NO in step S208), step S242.
Then, the normal control interval (t1) is set. As described above, according to the "large amplitude input control" of the present system ,: During normal traveling (NO in step S202), the vertical speed VG (that is, vertical acceleration) of the vehicle body is a predetermined threshold value (GA).
If it exceeds (YES in step S220), the damper force of the rear wheels is made hard (step S6 in FIG. 13).
0). Further, the upper limit value PLMT of the damping force is expanded. : On the other hand, when a large G input is made during turning (YES in step S202), etc. (Y in step S208)
In (ES), increasing the damping force too rapidly affects the maneuverability, so the response speed at which the damping force is hardened is slowed in step S212. G-through control The G-through control is to softly change the damper characteristic so that the fluctuation component included in the vertical acceleration signal G is not reflected in the riding comfort as it is during running on a rough road.

Details of this G-through control are shown in FIG. In step S250 of FIG. 33, if the value of the flag F for checking the value of the flag F is 2, the G-through control is not performed and the process returns to the main routine. When F = 2, the process proceeds to step S252 and thereafter. That is, the large-amplitude input control described above is a control for making the damper hard.
Since the through control is a control for changing the damper softly, the two controls are mutually exclusive depending on the value of the flag F so that the two controls do not interfere with each other. Further, due to the existence of step S250, the large-amplitude input control has a higher priority than the G-through control. This is because safety is prioritized over riding comfort in this embodiment.

A case where the large amplitude input control is not performed will be described. In this case, the process proceeds to step S252 and thereafter,
In steps S252 to S256, the threshold correction coefficients G0, G1 and G2 are calculated, and in step S258 the final threshold GTR is calculated as GTR = G0.G1.G2 (17). In step S260, this threshold value is compared with the vertical acceleration G to determine whether or not a large acceleration input has been made. G0 is determined based on the vehicle speed VB according to the characteristic shown in FIG. 34, G1 is determined based on the steering angle θH according to the characteristic shown in FIG. 35, and G2 is determined based on the steering angular velocity θ′H shown in FIG. 36. Determined according to characteristics.

Here, the acceleration G in step S258 is the average value of the output signals from the three acceleration sensors in the first and second embodiments, or G is the maximum value thereof. You may do it. When a large acceleration is input, the process proceeds to step S262 and “G
The flag F is set to 3 to indicate that the "through control" is executed. In step S264, it is determined whether slalom traveling is being performed. This determination can be made based on, for example, the amount of change in the steering wheel steering angle θH per unit time. If it is determined that the vehicle is traveling in slalom mode, the upper limit value PLMT is set in step S278.
To 1.2 times the normal time. This is because when the slalom traveling is performed, the damping force in the hard direction can be greatly changed to maintain the stability of the vehicle body. When the slalom traveling is not performed, step S5 in FIG.
In FIG. 6, PRTRn is attenuated by one step (soft).
This is to ensure a comfortable ride by prohibiting the damper force from being significantly changed in the hard direction. Also,
When it is determined that the slalom traveling is not performed (step S264), the upper limit value PLMT is reduced to 0.8 times the normal value in step S266.

If it is determined in step S264 that slalom traveling is not being performed, the control interval tx is changed according to the value of the lateral acceleration G. That is, when it is determined that the lateral direction G is larger than the threshold lateral G0 (| horizontal G | ≧ horizontal G0), a short control interval (t0) is set in step S270, and the lateral direction G is the lateral threshold value. If it is judged that it is smaller than G0 (| horizontal G | <horizontal G0),
In step S282, a longer control interval (t1) is set. However, t1> t2. The control in steps S272 to S276 is the same as step S in the above-mentioned "large amplitude input control".
The same as 214 to step S218, that is, the target damping force is between the front wheel and the rear wheel, or between the right wheel and the left wheel,
Make sure there is no difference of three steps or more.

On the other hand, if it is determined in step S268 that the lateral direction G is smaller than the threshold lateral G0 (| horizontal G | <horizontal G0), the normal control interval (t1) is set in step S282. . Thus, according to this "G through control" :: The vertical acceleration (that is, vertical speed) of the vehicle body is a predetermined value GTR.
If it is larger than, the flag F is set to 3 in step S262.
By setting to, the damping force is corrected in the soft direction in step S56. Also, by reducing the upper limit value PLMT, excessive input is blocked. : However, during slalom, the upper limit is expanded to allow a large change in the soft direction to be set. : Also, acceleration is generated in the lateral direction (step S
268), the softening of the damping force is delayed by increasing the control interval. Further, the running stability is improved by preventing the difference in the damping force between the forest and the feeling of the left and right wheels from increasing. : By giving priority to "large amplitude input control" over "G through control", priority is given to safety.

[0057]

As described above, according to the present invention,
It is possible to provide a vehicle suspension device that secures both low cost and suspension characteristic control at a high level.

[Brief description of drawings]

FIG. 1 is a diagram conceptually showing a relationship between a signal input to a suspension control device of an embodiment and control means.

FIG. 2 SD used in the suspension device of the embodiment
The figure which shows the characteristic of a damper.

FIG. 3 is an AD used in the suspension device of the embodiment.
The figure which shows the characteristic of a damper.

FIG. 4 is a table diagram showing priorities among various controls applied to the suspension control system of the embodiment.

FIG. 5 is a map diagram showing an application area between various controls applied to the suspension control system of the embodiment.

FIG. 6 is a block diagram showing the relationship between various controls applied to the suspension control system of the embodiment.

FIG. 7 is a diagram showing a relationship between actuators and wheel positions in the suspension device according to the first embodiment.

FIG. 8 is a diagram showing the relationship between various signals, various controls, and actuators in the suspension device according to the first example.

FIG. 9 is a flowchart for front wheel damping force control according to the first embodiment.

FIG. 10 is a graph showing the characteristic of the threshold value VG0 with respect to the vehicle speed VB.

FIG. 11 is a graph showing the characteristic of the coefficient K1 with respect to the vehicle speed VB.

FIG. 12 is a graph showing a characteristic of an upper limit value PLMT of damping force with respect to a vehicle speed VB.

FIG. 13 is a flowchart for rear wheel damping force control according to the first embodiment.

FIG. 14 is a graph showing a characteristic of a rear wheel target damping force PR defined by a vehicle speed VB in the first embodiment.

FIG. 15 is a graph showing a characteristic of a rear wheel target damping force PRTR defined by the front wheel damping force POF in the first embodiment.

FIG. 16 is a diagram showing a relationship between actuators and wheel positions in the suspension device according to the second embodiment.

FIG. 17 is a diagram showing the relationship between various signals, various controls, and actuators in the suspension device according to the second example.

FIG. 18 is a flowchart for rear wheel damping force control according to the second embodiment.

FIG. 19 is a flowchart for front wheel damping force control according to the second embodiment.

FIG. 20 is a graph showing the characteristics of the target damping force PF of the front wheels defined by the vehicle speed VB in the second embodiment.

FIG. 21 is a graph showing the characteristics of the target damping force PFTR of the front wheels defined by the rear wheel damping force POR in the second embodiment.

FIG. 22 is a diagram showing a relationship between actuators and wheel positions in the suspension device according to the third example.

FIG. 23 is a diagram showing the relationship between various signals, various controls, and actuators in the suspension device according to the third example.

FIG. 24 is a flowchart for damping force control during bounce and pitch control according to the third embodiment.

FIG. 25 is a flowchart for damping force control during roll control according to the third embodiment.

FIG. 26 is a graph showing the characteristic of the front wheel target damping force PF with respect to the steering angle θH in the third embodiment.

FIG. 27 is a steering angle change θ′H of the coefficient A in the third embodiment.
The graph which shows the characteristic with respect to.

FIG. 28 is a graph showing the characteristic of the coefficient K with respect to the vehicle speed VB in the third embodiment.

FIG. 29 is a flowchart of large-amplitude input control used in the suspension devices of the first to third embodiments.

FIG. 30: Coefficient GA used in large amplitude input control
The graph figure which shows the characteristic with respect to the vehicle speed VB.

FIG. 31: Coefficient GB used in large amplitude input control
FIG. 6 is a graph showing the characteristics of the steering angle θH.

FIG. 32 is a diagram schematically illustrating the operation of large amplitude input control.

FIG. 33 is a flowchart of G-through control used in the suspension devices of the first to third embodiments.

FIG. 34 is a graph showing characteristics of a coefficient G0 used for G-through control.

FIG. 35 is a graph showing the characteristics of the coefficient G1 used for G-through control.

FIG. 36 is a graph showing the characteristics of the coefficient G2 used for G-through control.

[Explanation of symbols]

1FL, 1FR ... High-speed motor, 1RL, 1RR ... Low-speed motor,
1'FL, 1'FR ... Low speed motor, 1'RL, 1'RR ... High speed motor, 2L, 2R, 2'L, 2'R ... Vertical G sensor

Claims (5)

[Claims] 1. A vehicle suspension device comprising a damper having a variable damping force on wheels, wherein a first vertical acceleration sensor is provided at substantially the center of a substantially front portion of the vehicle, and a substantially center of a substantially rear portion of the vehicle. A second vertical acceleration sensor, a steering angle detecting means for detecting a steering angle of the vehicle, and a bounce and pitch of the vehicle based on output signals of the first and second vertical acceleration sensors. A vehicle suspension device comprising: a control unit that controls the movement by feedback control and that controls the turning of the vehicle by feedforward control in response to a steering angle signal from the detection unit. 2. The vehicle suspension device according to claim 1, further comprising vehicle speed detecting means for detecting a vehicle speed, wherein the control means includes a vehicle speed signal from the vehicle speed detecting means and a steering angle from the steering angle detecting means. A vehicle suspension device characterized in that a vehicle turning control is performed in a feedforward manner based on a signal. 3. The turning control according to claim 1, wherein the turning control is performed when an operating condition for performing the turning control and an operating condition for bounce / pitch control are detected for the same wheel. A suspension device for a vehicle, characterized in that 4. The vehicle suspension device according to claim 1, further comprising means for detecting a straight traveling state and a turning state, and the turning control should be performed when the turning detection means detects the turning state of the vehicle. A vehicle suspension device characterized in that it is judged to be a driving condition, and when the turning detection means detects a straight traveling state of the vehicle, it is judged to be a driving condition for performing the bounce / pitch control. 5. The vehicle suspension device according to claim 1, wherein the turning control by the control means increases the damping force by a predetermined amount with respect to the current damping force.