JP3167835B2 - Vehicle suspension device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

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

[0002]

2. Description of the Related Art Conventionally, as a suspension device for a vehicle, for example, as disclosed in Japanese Patent Application Laid-Open No. 3-182826, a fluid cylinder is disposed between a vehicle body and each wheel. So-called full-active control suspension device (AC) that independently controls the supply and discharge of each wheel by a flow control valve to vary the suspension characteristics of the vehicle according to the driving state
S device) is known.

[0003] However, this full-active control suspension device has the disadvantage that the system becomes large-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 variable damping force damper capable of changing the damping force.
Has been proposed. This variable damper uses a fluid cylinder whose interior is separated into two chambers, and communicates between the two chambers with an orifice to control the throttle amount by the orifice. The control of the orifice throttle amount is performed as follows. That is, for example, a plurality of holes are provided in two disks,
The two disks are overlapped with their centers aligned, and located at the boundary between the two chambers in the cylinder. Then, one disk is fixed, the other disk is rotated by a stepping motor or the like, and an orifice is formed by overlapping the hole provided in the one disk with the hole provided in the other disk. Will be. The holes provided in the one disk have a plurality of sizes, and the holes provided in the other disk also have a plurality of sizes. Therefore, the rotation position of the step motor indicates the amount of throttle, that is, the damping force.

[0004]

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

However, the conventional suspension device necessarily requires three vertical acceleration sensors for performing bounce control, pitch control, and roll control. This has led to an increase in the cost of the suspension device. SUMMARY OF THE INVENTION An object of the present invention is to provide a vehicular suspension device that ensures a balance between low cost and suspension characteristic control at a high level.

[0006]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention relates to a vehicle suspension device having a damper having a variable damping force on a wheel. A first vertical acceleration sensor, a second vertical acceleration sensor provided substantially at the center of a substantially rear portion of the vehicle, steering angle detecting means for detecting a steering angle of the vehicle, the first and second vertical acceleration sensors. Control means for performing control for preventing bounce and pitch movement of the vehicle based on an output signal of the directional acceleration sensor by feedback control, and performing turning control of the vehicle by feedforward control in response to a steering angle signal of the detection means; It is characterized by having.

[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 the acceleration sensors can be reduced, the provided sensors are arranged in the longitudinal direction of the vehicle body, so that the bounce control and the pitch control can be sufficiently performed by the signals from these two acceleration sensors. In addition, the roll motion is substituted by turning control by feedforward control based on a signal from a steering angle sensor.

[0008]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, three preferred embodiments (first to third embodiments) to which the present invention is applied will be described. <Summary> In the suspension device of the above embodiment, a damper having the characteristics shown in FIG. 2 and a damper having the characteristics shown in FIG. 3 are used. The damper characteristic in FIG. 2 can change the damping characteristic independently in both the extending direction and the contracting direction. Therefore, this damper has a small number of stages (27 stages in the embodiment).
Can be applied to the damping characteristic control according to the so-called “Skyhook model”. Hereinafter, such a damper is referred to as an “SD” (sky hook damper) damper. In the damper characteristics shown in FIG. 3, if the damping characteristics in the extension direction are controlled, the characteristics in the contraction direction also change, so that only a coarse number of stages (five stages 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 stages at high speed. On the other hand, since the AD damper has only five steps, a low-speed step motor (actuator) is sufficient.

FIG. 1 shows what signals are commonly input to execute suspension characteristic control in the above three embodiments. That is, the control means of these embodiments comprises a vehicle speed signal VB and a steering wheel angle signal θH.
, A vertical acceleration signal G, and a brake signal Br. The control means receives these signals and
Either “sky hook 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 left and right front wheels, and AD dampers are used for left and right rear wheels. In the second embodiment, as shown in FIGS. 16 and 17, AD dampers are used for left and right front wheels, and SD dampers are used for left and right rear wheels.
A damper is used. In the first and second embodiments, high-speed S
A high speed feedback control using the skyhook model is used for the D damper, and the vehicle speed VB,
Since the feedforward control using the steering angle .theta.H and the like as a parameter is used, high-performance suspension performance can be secured although it is inexpensive. In particular, the first
In the embodiment, the front suspension aims to secure the posture stability (5 db in a low-frequency region) and the riding comfort, and the rear suspension also aims to secure the riding comfort. : In the third embodiment, as shown in FIGS. 22 and 23, SD dampers are used for the left and right front wheels and the left and right rear wheels. In the third embodiment, a pitch motion, a bounce motion,
Of the roll motion, 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). This eliminates the need for an acceleration sensor 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) is to correct the damper characteristic in the soft direction to secure the ride comfort and perform such vertical G movement when there is a relatively large amplitude vertical G movement. If the movement is appropriate during high-speed driving or turning, the correction in the soft direction is limited to ensure steering stability. -2: "Large-amplitude input control" (FIG. 29) increases the damping force in the case where there is an up-and-down G motion with a larger amplitude, with first consideration of ensuring steering stability. When such a vertical G motion matches during high-speed driving or turning, the correction in the hard direction is further increased to secure further steering stability. -3: "G through control" and "large amplitude input control"
Since the vertical G motion is a problem, they may interfere with each other. However, in this embodiment, there is no problem because “large amplitude input control” is given a higher priority than “G through control”. <Damper> FIGS. 2 and 3 show SD damper characteristics and AD, respectively.
Shows damper characteristics. The horizontal axis indicates the rotational position P of the step motor, and the vertical axis indicates the damping force. As described above, 27 positions are set in the SD damper, and 5 positions are set in the AD damper.

In the SD damper of FIG. 2, when the rotational position P moves in the forward direction, the damping force in the extension direction increases, but the damping force in the contraction direction becomes small. If the rotational position P is at a negative position, the damping force in the extension direction becomes small but the damping force in the contraction direction increases. That is, if the step motor is in a more positive position, the
The so-called "hard" characteristic is exerted by the damping force in the elongation direction, but the so-called "soft" characteristic is obtained for the movement of lowering the vehicle body. Also, if the stepping motor is at a more negative position, the so-called "hard" characteristic is exerted by the damping force in the contraction direction for the motion of lowering the vehicle body, but the vehicle body is raised. It has a so-called "soft" characteristic for exercise.

On the other hand, the AD damper is, as shown in FIG.
As the rotational position P of the step motor increases,
A "hard" characteristic is obtained in both the elongation direction and the contraction direction, and a "soft" characteristic is obtained as the rotational position becomes smaller. <Overall configuration of control system> Commonly used in the first to third embodiments is "steering control" relating to steering stability, and hardening of suspension characteristics when a large amplitude vertical acceleration is detected. "Large input amplitude control" to increase safety, and "G through" to make such acceleration "through" by softening suspension characteristics when detecting high-frequency vertical acceleration (acceleration when traveling on a rough road) Control "and" small amplitude control "for relatively softening suspension characteristics when a small amplitude vertical acceleration is detected. FIG. 4 and FIG. 5 are maps schematically showing the relationship between these control priorities and operation areas. 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 between 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, reference numerals 2L and 2R denote 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. The left and right SD dampers are driven by high-speed step motors 1FL and 1FR, respectively, and the left and right AD dampers for the rear wheels are driven by low-speed step motors 1R and 1FR, respectively.
L, 1RR is driven. The steering angle θH is detected by the steering angle sensor 3 and sent to the controller 10. The vehicle speed sensor 4
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, a steering control (steering control) section SH using the skyhook model includes acceleration signals GL and GR, a steering angle signal θH, a brake signal BR, and a vehicle speed signal V.
Input B or the like 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 a control procedure according to the flowchart of FIG. Also, the two damping force switching control units AA
A steering angle signal θH, a brake signal BR, and a vehicle speed signal VB are inputted to control two AD dampers for left and right rear wheels. Control unit A
The actual operation of A is realized by the controller 10 executing a control procedure according to the flowchart of FIG. As described above, since the feedback control that requires a high-speed operation is not applied to the rear wheel control, an expensive vertical G sensor is not required on the rear wheel side, and an inexpensive low-speed AD damper is used. Is sufficient.

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) performed based on the upper and lower G signals. Applied. Front wheel suspension control (SH / operation control): First implementation
Example The control procedure of the control unit SH is shown in FIG. The operation 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, and vehicle speed signal VB are input. In step S4, the acceleration signal G
L and GR are integrated, respectively, and the vertical speeds VGL and VG of the vehicle body are calculated.
Ask for R. In step S6, based on the vehicle speed VB, FIG.
The 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 it is determined in step S8 that | VGn | ≧ VG0, the flow proceeds to step S12. Conversely, if it is determined that | VGn | <VG0, in step S10, VGn is set to 0, whereby the vehicle body moves up and down. Is assumed not to move, and step S
Proceed to 12. As will be described later, since VGn is an important factor for determining the position of the damper, the region of the vehicle body vertical speed VGn where | VGn | <VG0 is a dead zone for control.

In step S12, a control gain K1 for determining the damper position P is calculated. The control gain K1 increases as the vehicle speed VB increases and as the steering wheel angular velocity (= time change) θ'H (= dθH / dt) increases. That is, when the vehicle speed VB or the steering angular speed θ'H is large, the SD damper position is changed at a high speed. In step S14, a target vehicle body vertical speed VGTRn that cancels the current vehicle body vertical movement speed (n is L
Left, R right), VGTRn = VGn · K1 (1) 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). Equation (2) changes the damping characteristic to one step harder so as to suppress the displacement of the vehicle body in the extension direction. Conversely, if VGTRn is negative, step S
The program proceeds to 18, where a target position PFTRn of the front wheel SD damper is calculated in accordance with PFTRn = PFn + 1 (3). Equation (3) is to change the damping characteristic to one step softly 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 control change tolerance as the vehicle speed increases. In step S24, the limit value PLMT is compared with the target position PFTRn, and the limit value is set to the target value PFTRn.
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 damper SD damper can achieve the target damping force. Here, the flag INHIBIT is set to “G through control” (step S276 in FIG. 33) or “large amplitude input control” (FIG.
The flag is set in step S218) of FIG. 9, and is set when there is a possibility that the front wheel suspension characteristics and the rear wheel suspension characteristics may be excessively different. Therefore, if this flag is set, step S30 is not executed and the motor position of the (front wheel) damper is not changed. Rear Wheel Suspension Control (First Embodiment) The suspension characteristics of the rear wheels are determined by five stages of AA dampers. In other words, the damping force of the damper is determined by the rotational positions of the motors 1RL and 1RR. FIG. 13 shows a control procedure of the motors 1RL and 1RR of the rear wheel AA damper.

The rear wheel suspension control of the first embodiment is as follows.
This is determined by feedforward control so that the rear wheels tend to be understeer with respect to the suspension characteristics of the front wheels determined by the control procedure of FIG. The reason for the feedforward control is that the control speed is high, and that the AA damper does not require advanced feedback control.

That is, in step S40, the vehicle speed VB
Is determined in accordance with 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 lines show the damping characteristics of the rear wheels while the vehicle speed is decreasing. When the vehicle speed is reduced, the characteristic is such that a lower damping force can be obtained compared to when the vehicle speed is increased. The characteristics shown in FIG. 14 are for obtaining a more understeer tendency in order to stabilize the vehicle body posture during deceleration.

In step S42, the current steering angle θH is compared with a predetermined threshold value θH0 to determine whether the vehicle is turning. When not turning (| θH | <θH0),
The process proceeds to step S52 and subsequent steps. In step S52, it is determined whether or not “large amplitude input control” is currently being executed.
In step S54, a flag (F = 3) is checked to determine whether or not "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.
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 flow proceeds to step S44. In step S44, the current motor position of the damper of the front wheel on the outside of the turn is monitored. This motor position is defined as POF. In step S46, FIG.
5, the target damping force of the rear wheel damper (ie,
Motor position PRTRn) is determined. The characteristic of FIG. 15 is such that a damping force lower than the damping force POF of the turning outer wheel obtained in step S44 is set on the rear wheel side. Since the front wheel side SD damper has 27 stages and the rear wheel side AA damper has 5 stages, the horizontal axis in FIG. 15 indicates the relative damping strength of the rear wheel damper with respect to the front wheel damper. That is, for example,
When the damping force of the front wheel is at a damping force corresponding to the fourth stage of the damping force of the rear wheel side damper (for example, the 25th stage), the damping force of the rear wheel side damper is one stage lower than the fourth stage. That is, it is reduced to three stages.

The characteristics shown in FIG. 15 may be changed so as to take the vehicle speed into consideration. In step S48, step S40
And the target damping force PR obtained according to the vehicle speed VB in step S4
In step 6, a comparison is made with the target damping force PRTRn obtained in relation to the damping force of the front wheels. If the latter is large (PR ≤ PRTR
n) In step S50, the target value of the rear wheel damping force is set to the damping force PR obtained in relation to the vehicle speed. That is, PRTRn = PR (4). This PRTRn is set to the rear step motor when step S30 in 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 for the motor in relation to the damper characteristics of the front wheels. The reason for this is that the rear wheel suspension control in the first embodiment is to ensure that the vehicle has understeer characteristics during turning. That is, for example, if the damping force PR determined according to the vehicle speed VB in step S40 is “third stage” and the damping force PRTRn determined in step S46 is “second stage”, PRTRn Since <PR, if PR is adopted as the damping force, the relationship that the rear wheels are understeered 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 ensured, the PR determined according to the vehicle speed VB as the rear wheel damping force in step S50 is adopted.

If it is determined in step S52 that "large amplitude input control" is being performed (F = 2), then in step S60,
Increase the damping force by one step. During "Large-amplitude input control" while traveling straight, the rear wheel side also has a damper characteristic on the hard side so that it can deal with large vertical movements of acceleration when riding on obstacles, rather than maintaining understeer characteristics. Because it is necessary to increase the In step S62, the control for correcting the characteristics of the rear wheels 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.

If the "G through control" is being executed while the vehicle is traveling straight (YES in step S54), the flow advances to step S54.
At 56, the rear wheel damper force is reduced by one step. The reason for lowering the suspension by one step is that, as described later, in the "G through control" (FIG. 33), the suspension characteristics of the front wheels are also changed to the soft side. 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 stages, and the driving of the damper is performed at a low speed step. Motor (1
(RL, 1RR), the cost can be reduced.
Since feedback control based on an acceleration signal requiring high-speed control is difficult to apply to such a low-speed damper, in the first embodiment, feed-forward control based on a vehicle speed signal VB (step S40 in FIG. 13) ). By adopting this feedforward control, the control can be simplified, which contributes to cost reduction. Although the feedforward control and the adoption of the low-speed damper may cause a decrease in the steering stability, the first embodiment employs the high-speed SD damper on the front wheel side and performs the skyhook control based on the vertical acceleration signal. (Step S12 in FIG. 9,
Since step S14) is performed, drivability can be ensured. : 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 at the time of turning affect the maneuverability, so that the damper characteristics of the rear wheels should be determined so that a predetermined relationship is established with the damper characteristics of the front wheels. In particular, in this embodiment, the damper characteristics of the rear wheels are determined such that the rear wheels maintain an understeer tendency with respect to the front wheels (FIG. 1).
Steps S48 and S50 of 3). : The setting of the control gain when determining the damper characteristics of the front and rear wheels increases as the vehicle speed VB increases, as the vehicle speed decreases (characteristic indicated by the broken line in FIG. 14), and as the steering angular velocity increases (K1 in FIG. 11). ), Set to be a hard trend. This is because the higher the vehicle speed VB, the more the vehicle is decelerated, and the higher the steering angular speed, the more the suspension characteristics need to be improved to improve the responsiveness. <Second Embodiment> The first embodiment employs an SD damper for the front wheels and an AD damper for the rear wheels. In the second embodiment, as shown in FIG. 16, an AD damper is provided on the front wheels,
An SD damper is used for the rear wheel. Therefore, in the second embodiment, the low-speed step motor 1'FL,
1'FR is provided, and a high-speed step motor 1'R is provided on the rear wheel.
L, 1'RR is provided. FIG. 17 shows a control block diagram of the second embodiment.

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

In step S60, the vertical acceleration signals GL and GR at the rear wheel position and the steering angle signal θ
H, various signals such as a brake signal BR and a vehicle speed signal VB are input. In step S62, the acceleration signals GL and GR are respectively integrated to obtain the vertical speed VG at the rear wheel position of the vehicle body.
Find L and VGR. In step S64, a speed threshold value VG0 is obtained based on the vehicle speed VB according to a characteristic diagram as shown in FIG. In step S66, the vertical speed VGn (n is L
Left and right) are compared with the threshold value VG0, respectively. If it is determined in step S66 that | VGn | ≥VG0, the process proceeds to step S70. Conversely, if it is determined that | VGn | <VG0, in step S68, VGn is set to 0, whereby the vehicle body moves up and down. Is determined not to have moved, and the process proceeds to step S70.

In step S70, a control gain K1 for determining the damper position P is calculated. The control gain is such that as the vehicle speed VB increases, the steering wheel angular velocity θ′H (= dθ
H / dt) indicates a larger value. In step S72, a target vehicle body vertical speed VGTRn that cancels the current vertical moving speed of the vehicle body is calculated as follows: VGTRn = VGn · K1 (5) If this VGTRn is positive, step S76
The target position PRTRn of the rear wheel SD damper is calculated according to the following equation: PRTRn = PRn-1 (6). Equation (6) changes the damping characteristic to one step harder so as to suppress displacement of the vehicle body in the extension direction. Conversely, if VGTRn is negative, step S
Proceeding to 76, the target position PRTRn of the rear wheel SD damper is calculated according to the following equation: PRTRn = PRn + 1 (7) The equation (7) is to change the damping characteristic by one stage to soft 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, the 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, 1'FR of the rear wheel damper are rotated so that the target damping force of the front wheel SD damper can be achieved. Here, the flag INHIBIT is a flag set in "G through control" and "large amplitude input control" as in the first embodiment. 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 become excessively different from each other, 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 five stages of AA dampers. In other words, the damping force of the damper is determined by the rotational positions of the motors 1'RL, 1'RR. FIG. 19 shows a control procedure of the motors 1'RL and 1'RR of the rear AA damper.

The front wheel suspension control of the second embodiment is as follows.
With respect to the front wheel suspension characteristics, the front wheel suspension characteristics are determined by feedforward control so that the rear wheel suspension characteristics determined by the control procedure of FIG. 18 tend to be understeer. 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 as shown in FIG. Here, in the characteristic diagram of FIG. 20, the solid line indicates the damping characteristic of the front wheels while the vehicle speed VB is increasing, and the broken line indicates the damping characteristics of the front wheels while the vehicle speed is decreasing. When the vehicle speed is reduced, the characteristic is such that a lower damping force can be obtained compared to when the vehicle speed is increased. The characteristics shown in FIG. 20 are for obtaining a more understeer tendency in order to stabilize the vehicle body posture during deceleration.

In step S102, the current steering angle θH is set to a predetermined threshold value θH0 in order to determine whether the vehicle is turning.
Compare with If the vehicle is not turning (| θH | <θH0), the process proceeds to step S102 and subsequent steps. Step S102 checks the flag (F = 2) to determine whether or not “large amplitude input control” is currently being executed. Step S104 is to check whether or not “G through control” is currently being executed (F = 2).
Check 3). If neither “large amplitude input control” nor “G through control” is executed, step S104
Then, when the process returns to the main routine, when the step S88 in FIG. 18 is executed, the target position PF set in the step S100 is changed to the motor 1'FL, 1'F of the front wheel damper.
Set to R.

If the vehicle is turning (| θH | ≧ θH0), the flow proceeds to step S104. In step S104, the current motor positions of the dampers of the turning outer wheels of the rear wheels are monitored. This motor position is defined as POF. Step S
In 106, the target damping force of the front wheel damper (that is, the motor position PFTRn) is determined according to the characteristics of FIG. FIG.
The characteristic 1 is such 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
And the target damping force PF obtained according to the vehicle speed VB in step S0
In step 6, a comparison is made with the target damping force PFTRn obtained in relation to the damping force of the rear wheel. 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 to the step motor for the front wheels when step S880 in FIG. 18 is executed. On the other hand, if it is determined in step S108 that PF ≦ PFTRn, the process proceeds to step S10 in relation to the damper characteristics of the rear wheels.
The target value PFTRn obtained in 6 is set for the motor.

In the front wheel suspension control of the second embodiment, the suspension characteristics of the front wheels are set such that the damping characteristics of the rear wheels are understeered compared to the damping characteristics of the front wheels during turning. That is, for example, if the damping force PF determined according to the vehicle speed VB in step S100 is “second stage” and the damping force PFTRn determined in step S06 is “third stage”, PFTRn > PF, if PF is adopted as the damping force, the relationship that the rear wheels are understeered 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 according to the vehicle speed VB as the rear wheel damping force in step S110 is adopted.

When it is determined in step S02 that "large amplitude input control" is being performed (F = 2), step S120 is performed.
Then, the damping force is increased by one step. `` Large amplitude input control '' while traveling straight
This is because it is necessary to enhance the damper characteristic on the rear wheel side to the hard side so that it can cope with a large acceleration vertical movement 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 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.

If the "G-through control" is being executed during straight traveling (YES in step S104), the rear wheel damper force is reduced by one step in step S106. The reason for lowering the suspension by one step is that, as described later, in the "G through control" (FIG. 33), the suspension characteristics of the front wheels are also changed to the soft side. <Third Embodiment> In the third embodiment, an SD damper is used for the front wheel and the rear wheel, and the detection in the vertical direction G is performed by one G sensor 6 provided on the front wheel side and the G sensor provided on the rear wheel side. This is performed by one G sensor 7. As shown in FIG. 22, if only one sensor is used for each of the front wheel side and the rear wheel side, it becomes difficult to detect the roll motion of the vehicle body. However, roll (steering) control is most necessary in the first place at the time of turning, and when turning, for example, if the front wheel characteristics are hardened, necessary and sufficient steering characteristics can be obtained. Conventionally, three or more sensors (a pair of sensors in the left and right direction, and one sensor before or after) are required. However, in the third embodiment, two sensors are sufficient, and the cost is reduced. It is useful for

FIG. 23 shows the overall configuration of the control system of the third embodiment. Roll (steering) control is based on the steering angle signal θH
The bounce and pitch components of the vehicle body posture are performed based on the vehicle speed signal VB and acceleration signals from the front and rear G sensors. Bounce and Pitch Control (Third Embodiment) FIG. 24 is a control flowchart for a bounce and pitch control portion 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 on the SD dampers of all the wheels.

In FIG. 24, steps S132 to S160 are substantially the same as steps S2 to S30 in FIG. 9, and the difference is that in the third embodiment, in step S132, the signal from the front G sensor is used. 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 motion speed VF of the front portion of the vehicle body and the vertical motion speed VR of the rear portion of the vehicle are input. A significant 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 a control procedure of roll control (steering control) performed on the SD dampers of the front and rear wheels. In step S170, it is determined whether the vehicle body is turning (| θH | ≧ θH0) or traveling straight (| θH | <θH0). If the vehicle is traveling straight (NO in step S170), a time change θ′H (= dθ / dt) of the steering angle is checked in step S192. It is determined that the vehicle is turning (YES) in step S170, or the steering angle has changed in step S192 (N
If it is determined as 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.
The reason for this is that, as described above, roll (steering) control is most necessary in the first place when turning, and when turning, for example, by making the front wheel characteristics hard, necessary and sufficient steering characteristics are obtained. Because you can do it.

The case where it is determined that the vehicle is turning or turning is described. In such a case, step S172 →
Proceeding to step S174, in step S174, the target damping force PF for the front wheels is changed to the vehicle speed VB and the steering angle θH.
Is determined based on The target damping force PF is, for example, as shown in FIG.
Are determined based on the vehicle speed VB and the steering angle θH according to the characteristics shown in FIG. That is, the characteristic shown in the drawing is such that the higher the steering angle θH and the higher the vehicle speed VB, the greater the damping force (hardening the damper characteristic). In step S178, the coefficient A is determined based on the steering angular velocity θ′H. The coefficient A has such a characteristic that it increases as the steering angular velocity θ′H increases, as shown in FIG. 27, for example. Step S17
In step 8, a 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 be larger as the vehicle speed is higher, the steering angle is larger, and the steering angular speed is larger. In step S180, a coefficient K for the rear wheel is determined. For example, as shown in FIG. 28, the coefficient K is a coefficient smaller than 1, and has a feature that the coefficient K decreases as the vehicle speed VB increases.

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

On the other hand. If the current damping force PRn of the rear wheel is smaller than the target damping force PFAn of the front wheel in step S182, the rear wheel may be over-steered. In step S188, PFTRn = P + -... ( 12) The meaning of the equation (12) is that the damping force PFTRn of the front wheel is set so that the front wheel on the outside of the turn has a hard characteristic in the contraction direction and the front wheel on the inside of the turn has a hard characteristic in the extension direction. It is to do. In step S190, PRTRn = PRn · K (13) 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: the respective signals from the 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 to suppress bounce / pitch based on GF, GR and signal θH from steering angle sensor (FIG. 24)
And turning control (roll control) is performed in a feed-forward manner based on the steering angle signal θH. By doing so, it is possible to reduce the number of G sensors by one compared to the related art, and still realize both the bounce / pitch control and the roll control. The feed forward control for the roll control is corrected for the front wheels by the vehicle speed VB and the steering angular speed θ′H (steps S174 and S178). : First, the damper force for the front wheels is determined, and thereafter, the damper characteristics of the rear wheels are determined to be understeer characteristics than the front wheels (step S182 to step S182).
190). : Roll control (turn control) is given priority over bounce / pitch control by using flag F.
Thereby, posture control in the roll direction at the time of turning is ensured. : Roll control is performed by feedforward control only during turning. <Large Amplitude Input Control and G Through Control> The three embodiments (FIGS. 8, 17, and 23) have been described. Next, a large amplitude input control and a 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, for example, in order to ensure safety when the vehicle body rides on an obstacle or the like.
Is a control for detecting that the input has a large amplitude and increasing the damping force.

FIG. 29 shows a control procedure of the large amplitude input control. The arbitration between the large amplitude input control (FIG. 29) and, for example, the SH control in the first embodiment 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, the steps from step S60 are executed.

First, with reference to the flowchart of FIG. 29, what kind of control is performed 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 is compared with a threshold value θH0 to determine whether the vehicle is turning. The control differs 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 is, for example, a time interval in which the control procedure of FIG. 13 of the first embodiment is executed. The shorter the time interval, the sooner the control is performed, and the more sensitive it is to the input. In the control procedure of FIG. 29, the control interval "slow" means that the control interval tx is set to tx = t0 when t0>t1> t2, and the "faster" means that the control interval tx is set to tx = t2. , "Normal" means that tx = t1 is set. In FIG. 32, "enlarging" the upper limit value PLMT means expanding the characteristic of FIG. 12 further by 1.2 times.

If the vehicle is traveling straight (| θH | <θH0), the flow advances to step S219 to determine a threshold value GA for determining whether or not a large amplitude input has been made. This threshold value GA is determined based on the vehicle speed VB according to, for example, a characteristic as shown in FIG. The characteristics of the threshold value GA in FIG. 30 are such that the higher the vehicle speed VB, the higher the threshold value. If the vertical vehicle body speed VG is smaller than GA, that is, if | VG | <GA, the process proceeds to step S230, where 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 the value of "normal" (FIG. 12) is set because the PLMT is not changed.

When it is determined in step S220 that the vehicle speed VG is higher 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 expanded to 1.2 times. On the other hand, when there is a large amplitude input during the straight traveling, it is determined that the vehicle is traveling straight in step S42 of the control procedure of FIG. Can 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 is performed.
It will go to. In step S60, to make the target damping force 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 the continuation is that such a large amplitude input state is considered to continue for the predetermined time.

As described above, when a large amplitude is input during straight traveling, the damper force control for the rear wheels (FIG. 13) and the "large amplitude input control" (FIG. 29) operate in cooperation. To deal with. That is, when there is a large amplitude input while traveling straight,
For the rear wheels, the damping force is increased (step S6).
0), by shortening the control interval (time interval t2) between the front wheel control (FIG. 9) and the rear wheel control (FIG. 13), the response to the shock input is sharpened. Further, in some cases, the damping force must be increased by accelerating the reaction. In such a case, the upper limit value PLMT is increased (step S226) so that the damping force is not clipped even if the damping force increases due to the enhancement of the damping force in response to the shock input.

When turning (| θH | ≧ θH in step S202)
When it is determined as 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. This threshold value GB is determined in step S206 based on the steering angle θH and the steering angular velocity θ′H, for example, according to a characteristic as shown in FIG. This characteristic is such that the larger the steering angle θH and the larger the steering angular velocity θ′H, the larger the value of GB.

If the impact input applied to the vehicle body is large such that | VG | ≧ GB, the flag F is set at 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, step S214, step S216
In the case of, the difference in the target damping force between the front wheel and the rear wheel or between the right wheel and the left wheel is prevented from being more than three steps. When a difference of three or more steps is about to occur between the front wheel and the rear wheel, | Pf−Pr | ≧ 3 (15), and a difference of three or more steps is about to occur between the right wheel and the left wheel. Then, | PL−PR | ≧ 3 (16) should be satisfied. Here, 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 output, for example, in step S1 of FIG.
Reference numeral 58 denotes a signal for controlling whether or not to actually output a damping force signal for each of the front wheel and rear wheel dampers. When the signal INHIBIT is generated, the change of the damping force is stopped, so that the generated damping force is applied between the front wheel and the rear wheel or between the right wheel and the right wheel.
There is no difference of three or more steps between the left wheels.

On the other hand, even when the vehicle is turning, if the impact force is small (NO in step S208), step S242 is executed.
At the normal control interval (t1). As described above, according to the "large-amplitude input control" of the present system: During normal running (NO in step S202), the vertical speed VG (that is, vertical acceleration) of the vehicle body is equal to a predetermined threshold (GA).
Is exceeded (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 also extended. : On the other hand, when a large G input is made during turning (YES in step S202) (Y in step S208).
In the case of ES), increasing the damping force too quickly affects the stability, so that the response speed for hardening the damping force is reduced in step S212. G Through Control The G through control is to change the damper characteristic to soft so that the fluctuation component included in the vertical acceleration signal G is not directly reflected in the riding comfort when traveling on a rough road or the like.

FIG. 33 shows the details of the G through control. In step S250 of FIG. 33, when the value of flag F for checking the value of flag F is 2, the process returns to the main routine without performing the G-through control. When F = 2, the process proceeds to step S252 and subsequent steps. That is, the above-described large-amplitude input control is a control for hardening the damper.
Since the through control is a control for changing the damper to soft, the two controls are mutually exclusive according to the value of the flag F so as not to interfere with each other. Also, due to the presence of step S250, the large-amplitude input control has a higher priority than the G-through control. This is because in this embodiment, safety was given priority over ride comfort.

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

Here, in the first and second embodiments, the acceleration G in step S258 is an average value of the output signals from the three acceleration sensors or a value indicating the maximum value thereof. You may do so. If a large acceleration has been input, the process proceeds to step S262, where "G
The flag F is set to 3 to indicate that the "through control" is executed. In step S264, it is determined whether the vehicle is running in slalom. This determination can be made, for example, based on the amount of change in the steering wheel angle θH per unit time. If it is determined that the vehicle is running in slalom, the upper limit PLMT is determined in step S278.
Is expanded to 1.2 times the normal time. This is because when the vehicle is running in slalom, the damping force in the hard direction can be largely changed to maintain the stability of the vehicle body. If the vehicle is not running slalom, step S5 in FIG.
At 6, PRTRn is attenuated (softened) by one stage.
This is because the riding comfort is ensured by prohibiting the damper force from being largely changed in the hard direction. Also,
If it is determined that slalom running is not being 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 running 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 horizontal direction G is larger than the threshold horizontal G0 (| horizontal G | ≧ horizontal G0), a short control interval (t0) is set in step S270, and the horizontal direction G is set to the threshold horizontal G0. If it is determined 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 that in the above-described "large amplitude input control".
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.
The difference of three or more steps should not occur.

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

[0057]

As described above, according to the present invention,
It is possible to provide a suspension apparatus for a vehicle in which both low-cost and low-cost suspension characteristic control are ensured at a high level.

[Brief description of the drawings]

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

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

FIG. 3 shows 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 a relationship between various controls applied to the suspension control system of the embodiment.

FIG. 7 is a diagram showing a relationship between an actuator and a wheel position in the suspension device according to the first embodiment.

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

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

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

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

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

FIG. 13 is a flowchart for controlling the damping force of the rear wheels according to the first embodiment.

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

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

FIG. 16 is a diagram showing a relationship between an actuator and a wheel position in the suspension device according to the second embodiment.

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

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

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

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

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

FIG. 22 is a view showing a relationship between an actuator and a wheel position in the suspension device according to the third embodiment.

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

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

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

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

FIG. 27 shows a steering angle change θ′H of a coefficient A in the third embodiment.
FIG.

FIG. 28 is a graph showing characteristics of a coefficient K with respect to a 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 shows a coefficient GA used in large-amplitude input control.
FIG. 4 is a graph showing characteristics of the vehicle speed with respect to a vehicle speed VB.

FIG. 31 shows a coefficient GB used in large-amplitude input control.
FIG. 6 is a graph showing characteristics of the steering angle with respect to 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 characteristics of a coefficient G1 used for G-through control.

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

[Description of Signs] 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

──────────────────────────────────────────────────続 き Continuation of the front page (56) References JP-A-5-221220 (JP, A) JP-A-5-24423 (JP, A) JP-A-4-81315 (JP, A) 110932 (JP, A) JP-A-5-155221 (JP, A) JP-A-5-155220 (JP, A) JP-A-5-104926 (JP, A) JP-A-4-27612 (JP, A) (58) Field surveyed (Int. Cl. ⁷ , DB name) B60G 1/00-25/00

Claims (5)

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