JPH0761221A - Suspension device for vehicle - Google Patents

Abstract

PURPOSE:To dispense with a weight sensor and to make cost reduction and the improvement of suspension characteristics compatible by reflecting a natural frequency which can be obtained through the frequency analysis of a signal based on upward speed and applied to the measurement of vehicle weight, to control characteristics. CONSTITUTION:A suspension device has a vertical acceleration sensor 2L, 2R provided on its right and left sides, and high speed stepping motors 1FL, 1FR to drive each of right and left sky hook dampers for a front wheel and low speed stepping motors 1RL, 1RR to drive right and left adaptive dampers for a rear wheel are controlled by an ECU 10. In the case of front wheel suspension control, the vertical speeds VGN (VGL, VGR) of a car body obtained by integrating acceleration signals is compared with the threshold VGO of speed based on a car speed VB. In the case of 1VGN>=VGO, a control gain K1 for determining a damper position P is computed for determining a target car body vertical speed to cancel the vertical moving speed of the car body at the present time, and the target position of a front wheel SD damper is thereby determined.

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, for example, a variable damper whose damping force can be changed by an actuator or a suspension whose hydraulic pressure control can change suspension characteristics. The present invention relates to a suspension device capable of determining a vehicle weight from a directional acceleration signal and providing suspension characteristics suitable for the vehicle weight at that time.

[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.

Since the full active suspension control system uses high 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 control device" only changes the damping force, so the vehicle height control (posture control) accuracy is lower than that of the full active suspension control device, but it is advantageous in cost. .

[0005]

The basic characteristics of the suspension are usually determined on the basis of the total weight at that time, assuming a state in which a driver having a standard weight is driving, for example. Is. However, for example, when more than usual occupants board a vehicle having a soft suspension, the bouncing motion may be reduced more than usual. This is because the total weight of the vehicle has increased and the natural frequency has decreased, so that there is no resonance with the natural frequency of the suspension.

As described above, the original suspension device should determine the vehicle weight at that time and set the suspension characteristics according to the weight. Conventionally, the vehicle weight is measured directly by mounting a load sensor on the vehicle body. However, the addition of such a sensor has led to an increase in cost of the suspension device.

The present invention is to provide a vehicle suspension device which ensures both low cost and suspension characteristic control at a high level.

[0008]

In order to achieve the above object, the present invention inputs a detected vertical acceleration signal,
In a vehicle suspension device that feedback-controls a suspension characteristic of a vehicle in a direction in which the value of the acceleration signal decreases, a detection unit that detects a vertical acceleration signal and a suspension characteristic of the vehicle in a direction in which the value of the acceleration signal decreases. Control means for feedback control,
Analyzing means for frequency-analyzing the detected vertical acceleration signal, determining means for determining the natural frequency of this vehicle from the analysis result, and means for changing the control characteristic of the control means based on the determined natural frequency. And is provided.

[0009]

Since the total weight of the vehicle is reflected in the vertical acceleration, the natural frequency obtained by frequency analysis of the signal is the basis for estimating the vehicle weight. Therefore,
If this estimated natural frequency is reflected in the control characteristics,
The cost can be reduced and more accurate suspension characteristics can be obtained.

[0010]

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 to execute the suspension characteristic control in the above-mentioned three embodiments. That is, the control means of these embodiments has a vehicle speed signal V _B and a steering wheel steering angle signal θ.
_H , 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 V _B , for the low-speed AD damper.
Since feed-forward control using the steering angle θ _H as a parameter is used, 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". : Determines the total weight of the vehicle and changes the suspension characteristics according to the total weight. With this change,
The suspension characteristics most suitable for the weight at that time can be obtained. -1: For the determination of the total vehicle weight, the vertical acceleration signal is subjected to frequency analysis, and the weight is determined based on the frequency component of the maximum amplitude. An acceleration sensor is indispensable for a suspension device, and the signal from this sensor is used for optimal suspension characteristics and also for determining the weight, which is also indispensable. Therefore, both cost reduction and suspension characteristics optimization can be achieved. it can. -2: The estimated gross weight data is sent to the ABS controller of the vehicle of this example. This is because the vehicle weight data is also important for ABS control in terms of control optimization. <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 V _B.

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, 2 _L and 2 _R are vertical acceleration sensors provided on the left and right, respectively, and generate acceleration signals G _L and G _R. This acceleration signal is sent to the controller 10. Further, the high speed step motors 1 _FL and 1 _FR drive the left and right SD dampers, respectively, and the low speed step motors 1 _RL and 1 _RR drive the left and right AD dampers for the rear wheels, respectively. The steering angle θ _H is detected by the steering angle sensor 3 and sent to the controller 10. The vehicle speed V _B detected by the vehicle speed sensor 4 and the signal B _R 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) unit SH using the skyhook model has acceleration signals G _L and G _R , a steering angle signal θ _H , a brake signal B _R , 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 θ _H , the brake signal B _R , and the vehicle speed signal V _B 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 signal G_L，
G_R, Steering angle signal θ_H, Brake signal B_R, Vehicle speed signal V_BEtc.
Input various signals. In step S4, the acceleration signal G
_L, G _R And the vertical velocity V of the vehicle body_GL, V
_GRAsk for. In step S6, the vehicle speed V_BFigure based on
Velocity threshold value V according to the characteristic diagram of 10_G0Ask for. Step
In step S8, the vertical velocity V_Gn(N indicates L left and R right
Respectively, the above-mentioned threshold V _G0 Compare with. Step S
8 with | V_Gn│ ≧ V_G0If it is determined that step S12
, But conversely | V_Gn│ <V_G0If it is determined that
In step S10, V_GnBy setting = 0,
The car body is considered not to move up and down, and
Go to step S12. As will be described later, V_GnIs the position of the damper
Is an important factor in determining_Gn│ <V_G0And
Vertical velocity V of the vehicle_GnIs the dead zone of control
Becomes

In step S12, the control gain K ₁ for determining the damper position P is calculated. The control gain K ₁ has a larger value as the vehicle speed V _B is higher and the steering wheel angular velocity (= time change) θ ′ _H (= dθ _H / dt) is larger. That is, when the vehicle speed V _B and the steering angular speed θ ′ _H are large, the SD damper position is changed at high speed. In step S14, a target vehicle body vertical speed V _GTRn (n is L is set to cancel the current vertical moving speed of the vehicle body).
Left, R shows right), and V _GTRn = V _Gn · K ₁ (1) is calculated. If this V _GTRn is positive, step S20.
To the target position P _FTRn (n is L
Left, R right) are calculated according to P _FTRn = P _Fn -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 V _GTRn is negative, step S
Proceeding to 18, the target position P _FTRn of the front wheel SD damper is calculated according to P _FTRn = P _Fn +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. P _Fn 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 P _LMT as shown in FIG. 12 is obtained based on the vehicle speed V _B. Limit value P _LMT
Has a tendency to increase as the vehicle speed V _B 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 P _LMT is compared with the target position P _FTRn, and this limit value is compared with the target value P _FTRn.
If is exceeded, the target value is clipped to this limit value in step S26.

[0020] At step S28, checks whether flag INHIBI _T is set. If this flag is not set, in step S30, the motors 1 _FL and 1 _FR are rotated so that the front wheel SD damper can achieve the target damping force. The flag INHIBI _T is "G through control" (step S276 in FIG. 33) and "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 position of the motors 1 _RL and 1 _RR . FIG. 13 shows a control procedure for the rear wheel AA damper motors 1 _RL and 1 _RR .

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 V _B
The target damping force P _R (V _B ) of the rear wheels according 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 V _B 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 θ _H is compared with a predetermined threshold value θ _H0 to determine whether or not the vehicle is turning. When not turning (| θ _H | <θ _H0 ),
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.
Is executed, the target position P _R set in step S40 is set to the rear wheel damper motors 1 _RL and 1 _RR .

If the vehicle is turning (| θ _H | ≧ θ _H0 ), 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 P _OF . In step S46, FIG.
According to the characteristics of 5, the target damping force of the rear wheel damper (that is,
The motor position P _RTRn ) is determined. The characteristic of FIG. 15 is such that a damping force lower than the damping force P _{OF 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 P _R obtained in accordance with the vehicle speed V _B and the step S4
In step 6, the target damping force _PRTRn obtained in relation to the damping force of the front wheels is compared. If the latter is large (P _R ≤
P _RTRn ), the target value of the rear wheel damping force is set to the damping force P _R obtained in relation to the vehicle speed in step S50. That is, P _RTRn = P _R (4) This P _RTRn is set to the step motor of the rear wheel at the time when step S30 of FIG. 9 is executed. On the other hand, if it is determined in step S48 that P _R > P _RTRn , the target value P _RTRn 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 P _R determined according to the vehicle speed V _B in step S40 is “third stage” and the damping force P _RTRn determined in step S46 is “second stage” Since P _RTRn <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 hold. Therefore, only when the P _R ≦ P _RTRn in step S48, the words, only if the relation of understeer is ensured, since adopting a P _R determined according to the vehicle speed V _B as the rear wheel damping force at step S50 is there.

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), step S
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} , 1 _RR ) is adopted, the cost can be reduced.
Since feedback control based on an acceleration signal that requires high speed control is difficult to apply to such a low speed damper, in the first embodiment, feed forward control based on the vehicle speed signal V _B (steps in FIG. 13). S40) 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 such that the vehicle speed V _B becomes higher, the vehicle speed is decelerated (characteristic of the broken line in FIG. 14), and the steering angular speed becomes larger (in FIG. K ₁ ), so that it tends to be hard. This is because the higher the vehicle speed V _B , 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 step motor _1'FL ,
1 _'FR are provided, high speed of the step motor 1 to the rear wheels' _RL, 1' _RR is provided. FIG. 17 shows a control block diagram of this second embodiment.

FIG. 18 shows suspension control of the rear wheels, and FIG. 19 shows suspension control of the front wheels. 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, various signals such as vertical acceleration signals G _L and G _{R at} the rear wheel position, a steering angle signal θ _H , a brake signal B _R and a vehicle speed signal V _B are input. In step S62, the acceleration signals G _L and G _R are integrated to obtain the vertical velocity V at the rear wheel position of the vehicle body.
_{Find GL} and V _GR . In step S64, a speed threshold value V _G0 is obtained based on the vehicle speed V _B according to the characteristic diagram shown in FIG. In step S66, the vertical velocity V _Gn (n indicates L left and R right) is compared with the threshold V _G0 . If | V _Gn | ≧ V _G0 is determined in step S66, the process proceeds to step S70. On the contrary, if | V _Gn | <V _G0 is determined, V _Gn = 0 is set in step S68. It is assumed that the vehicle body is not moving in the vertical direction, and the process proceeds to step S70.

In step S70, the control gain K ₁ for determining the damper position P is calculated. This control gain is such that as the vehicle speed V _B increases, the steering wheel steering angular velocity θ ′ _H (= dθ
_The larger _H / dt), the larger the value. In step S72, a target vehicle body vertical speed V _GTRn that cancels the current vertical moving speed of the vehicle body is _calculated as V _GTRn = V _Gn · K ₁ (5). If this V _GTRn is positive, step S76.
Then, the target position P _RTRn of the rear wheel SD damper is calculated according to P _RTRn = P _Rn -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 V _GTRn is negative, step S
Proceeding to 76, the target position P _RTRn of the rear wheel SD damper is calculated according to P _RTRn = P _Rn +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. P _Rn 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 P _LMT as shown in FIG. 12 is obtained based on the vehicle speed V _B. Step S
At 82, the limit value P _LMT is compared with the target position P _FTRn. If the target value P _RTRn 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, to rotate the motor 1 _'FL, 1' _FR of the rear wheel damper front wheels SD damper as target damping force can be achieved. 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 the damper is determined by the rotational position of the motor 1 _'RL, 1' _RR. FIG. 19 shows a control procedure of the motors 1 ′ _RL and 1 ′ _RR of the rear wheel AA damper.

The front wheel suspension control of the second embodiment is as follows.
Control procedure of FIG. 18 for suspension characteristics of front wheels
The suspension characteristics of the rear wheels determined by
Make sure that the front wheel suspension features
Is determined by feedforward control.
It That is, in step S100, the vehicle speed V_BAccording to
Target front wheel damping force P_F(V _B) As shown in FIG.
Determine according to the characteristics. Here, in the characteristic diagram of FIG.
The solid line is the vehicle speed V_BOf the front wheels while the
This is the damping characteristic, and the broken line is only when the vehicle speed is decreasing.
The damping characteristics of the front wheels are shown. When the vehicle speed is decelerating
In comparison, the characteristics are such that a lower damping force can be obtained.
There is. The characteristic of Fig. 20 is to stabilize the body posture during deceleration.
In order to obtain a more understeer tendency.

At step S102, whether or not the vehicle is turning.
To determine whether the current steering angle θ _HIs a predetermined threshold θ_H0
Compare with. Not turning (| θ_H│ <θ_H0)In case
Goes to step S102 and thereafter. Step S102
Flag whether or not "Large amplitude input control" is currently being executed
(F = 2), and step S104 is "G scan."
Flag (F =
3) is investigated. "Large amplitude input control"
If the "routing control" is not executed either, step S104
Returns to the main routine.
When step S88 is executed, the settings are made in step S100.
Specified target position P_FIs the front wheel damper motor 1 '_FL, 1 '
_FRIs set to.

If the vehicle is turning (| θ _H | ≧ θ _H0 ), 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 P _OF . Step S
At 106, the target damping force of the damper of the front wheel (that is, the motor position P _FTRn ) 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 P _{OF of the} outer turning 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
At vehicle speed V_BTarget damping force P determined according to_FAnd step S0
Target damping force P calculated in relation to rear wheel damping force in 6_FTRnWhen
To compare. If the latter is small (P_F>
P_FTRn), In step S110, set the target value of the front wheel damping force to the vehicle
Damping force P calculated in relation to speed_FAnd That is, P_FTRn= P_F … (8) This P_FTRnHowever, step S880 in FIG.
When set, the front wheel stepper motor is set.
It On the other hand, in step S108, P_F≤P_FTRnIs judged
If so, step S10 is performed in relation to the damper characteristics of the rear wheels.
Target value P obtained in 6 _FTRnIs set to 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 P _F determined according to the vehicle speed V _B in step S100 is “second stage” and the damping force P _FTRn determined in step S06 is “third stage” Since P _FTRn > P _F , if P _F is adopted as the damping force, the relationship that the rear wheel is understeer with respect to the front wheel may not hold. Therefore, only when P _R > P _RTRn in step S108, that is, only when the understeer relationship is ensured, the rear wheel damping force P _F determined in accordance with the vehicle speed V _B is adopted in step S110. is there.

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 θ _H
And performed based on the vehicle speed signal V _B, the vehicle body attitude bounce, the pitch component, the vehicle speed signal V _B, carried out on the basis of the acceleration signal from the front and rear G sensor. 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 in the third embodiment, the front G sensor is operated at step S132. The difference is that the signal G _F and the signal G _R from the rear G sensor 7 are input, and in step S134, the vertical movement speed V _{F of the} front part of the vehicle body and the vertical movement speed V _{R 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 (| θ _H | ≧ θ _H0 ) or going straight (| θ _H | <θ _H0 ). If it is traveling straight (in step S170 NO), examine the time variation theta _'H of the steering angle (= d [theta] / dt) in a 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.
The case where it is determined that the vehicle has started or is turning will be described. In such a case, step S172 → step S1
In step S174, the target damping force P _F for the front wheels is determined based on the vehicle speed V _B and the steering angle θ _H. The target damping force P _F is determined based on the vehicle speed V _B and the steering angle θ _H according to the characteristics shown in FIG. 26, for example.
That is, the characteristics shown in the figure are such that the higher the steering angle θ _{H and the} higher the vehicle speed V _B , the greater the damping force (hardening the damper characteristics). Also determines based step S178 the coefficients A to the steering angular velocity theta _'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. In step S178, the target damping force P _FAn (n represents right or left) is calculated.

P _FAn = P _F · A (9) Thus, in steps S174 to S178,
The target damping force P _FAn 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 V _B increases.

In step S182, the target damping force P _FAn for the front wheels and the actual present damping force P _{Rn for the} rear wheels are compared. The target damping force P _{FAn of the} front wheels is the current damping force P _{Rn of the} rear wheels.
If it is larger than P _FAn ≦ P _Rn , the process proceeds to step S184, and in order to set the target damping force P _FAn to the final target damping force P _FTRn of the front wheels, P _FTRn = P _FAn. (10) In step S186, the coefficient K obtained in step S180 is used so that the rear wheels have an understeer tendency with respect to the front wheels, and P _RTRn = P _FAn · K (11) . 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, In step S182, if the current rear wheel damping force P _Rn is smaller than the front wheel target damping force P _{FA n} , the rear wheels may be oversteered, so in step S188, P _FTRn = P + -... (12) The expression (12) means that the damping force P _FTRn of the front wheels is such 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. It is to set. Further, in step S190, P _RTRn = P _Rn · K (13) is set so that the rear wheel characteristics have a lower damping force than the current damping force so that an 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 G _F , G _R and the signal θ _H from the steering angle sensor (FIG. 24)
The turning control (roll control) is performed in a feedforward manner based on the steering angle signal θ _H. 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 V _B 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 body speed V _G in the vertical direction. In step S202, the steering angle θ _H and the threshold θ _H0 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 V _G.

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, the control interval is “slow”, when the control interval t _x is t ₀ > t ₁ > t ₂ , t _x = t ₀ is set, and “early” is It is assumed that t _x = t ₂ is set, and “normal” is set as t _x = t ₁ . In addition, in FIG. 32, “enlarging” the upper limit value P _LMT means that the characteristic of FIG.
Say to spread twice.

When the vehicle is traveling straight (| θ _H | <θ _H0 ), the routine proceeds to step S 219, where a threshold value G _A for determining whether or not there is a large amplitude input is determined. This threshold value G _A is determined based on the vehicle speed V _B according to the characteristic shown in FIG. 30, for example. The characteristic of the threshold G _{A in} FIG. 30 is a threshold that increases as the vehicle speed V _B increases. If the vertical vehicle speed V _G is smaller than G _A , that is, | V _G | <G _A , the process proceeds to step S230, the flag F is set to 0, and the control cycle time t _{x is set} in step S232.
Is set to the normal interval (t ₁ ), and P _LMT is not changed, so the value of “normal” (FIG. 12) is set.

When it is determined in step S220 that the vehicle body speed V _G is greater than the threshold value G _A , the flag F is set to 2 in step S222. Then, in step S224, the control interval is set to "early" (t _x = t ₂ ),
Extend the upper limit value P _{LM T} 1.2 times at step S226.
On the other hand, when there is a large amplitude input while traveling straight,
Step S42 of the control procedure of (rear wheel control of the first embodiment)
In step S52, the value of the flag F is checked. As described above, since the flag F is set to 2 in step S222 of FIG. 29, YES is determined in step S52, and step S52 is performed.
It will proceed to 60. In step S60, in _order to set the damping force target value P _RTRn of the rear wheel damper to a harder tendency than the current value, P _RTRn = P _Rn +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) short (time interval t ₂ ), 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, the upper limit value P _LMT is increased (step S226) so that 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 (| θ _H | ≧ θ in step S202)
_If it is determined to be ( _H0 ), the magnitude of the impact is measured by comparing the predetermined threshold value G _B with the vertical speed V _{G of the} vehicle body in step S208. This threshold value G _B is determined in step S206 based on the steering wheel steering angle θ _H and the steering angular velocity θ ′ _H according to the characteristic shown in FIG. 31, for example. This characteristic is such that the larger the steering wheel steering angle θ _{H and the} larger the steering angular velocity θ ′ _H , the larger the value of G _B.

When the impact input applied to the vehicle body is large such that | V _G | ≧ G _B , 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, | P _f −P _r | ≧ 3 (15), and the difference between the right wheel and the left wheel is 3 steps or more. If so, | P _L −P _R | ≧ 3 (16). However, Pf is the final target damping force P of the front wheels.
_FTRn 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 INHBIT 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, even when the vehicle is turning, if the impact force is small (NO in step S208), step S242.
Then, the normal control interval (t ₁ ) 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 V _{G of the} vehicle body (that is, vertical acceleration) is a predetermined threshold value (G _A )
If it exceeds (YES in step S220), the damper force of the rear wheels is made hard (step S6 in FIG. 13).
0). Furthermore, the upper limit value P _LMT 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 G ₀ , G ₁ and G ₂ are calculated, and in step S258 the final threshold G _TR is calculated as G _TR = G ₀ · G ₁ · G ₂ (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. G ₀ is determined based on the vehicle speed V _B according to the characteristic as shown in FIG. 34, G ₁ is determined based on the steering angle θ _H according to the characteristic as shown in FIG. 35, and G ₂ is determined based on the steering angular velocity θ ′ _H. For example, it is determined according to the characteristics as shown in FIG.

Here, the acceleration G in step S258 and
Are three acceleration sensors in the first and second embodiments.
The average value of the output signals from, or their maximum value
You may make it show G as. Large acceleration
When there is an input of “G”, the process proceeds to step S262, and “G
The flag F is set to indicate that "through control" is executed.
Set to 3. In step S264, the slalom running is performed.
Determine if This judgment is made, for example, by the handle
Rudder angle θ_HJudgment based on the amount of change per unit time
be able to. Judged to be running in slalom
If it is found, the upper limit value P is determined in step S278. _LMT
To 1.2 times the normal time. Run slalom
If there is a large change in the damping force in the hard direction
This is because it is possible to maintain the stability of the vehicle body. Slallow
If the vehicle is not traveling, step S in FIG.
At 56, P_RTRnIs attenuated by one step (soft)
It Don't allow the damper force to change significantly in the hard direction
This is because the ride comfort is secured by stopping. Well
In addition, it was determined that he was not running slalom (
If step S264), then in step S266
Upper limit P_LMTIs reduced to 0.8 times the normal value.

If it is determined in step S264 that slalom traveling is not being performed, the control interval t _x 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 G ₀ (| horizontal G | ≧ horizontal G ₀ ), a short control interval (t ₀ ) is set in step S270, and the lateral direction G is set. Is smaller than the threshold horizontal G ₀ (| horizontal G | <horizontal G ₀ ),
Setting larger control interval (t ₁₎ at step S282. However, t ₁ > t ₂ . 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, when it is determined in step S268 that the lateral direction G is smaller than the threshold lateral G ₀ (| horizontal G | <horizontal G ₀ ), the normal control interval (t ₁ ) is determined in step S282. To set. Thus, according to this "G through control" :: The vertical acceleration (that is, the vertical speed) of the vehicle body is the predetermined value G _TR.
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 P _LMT , 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. <Change of control characteristic according to weight change> As described above,
Suspension characteristics are greatly affected by the total weight of the vehicle. The suspension device described below is an improvement of the suspension devices of the above-described first to third embodiments. The improvement is based on the determination of the total weight of the vehicle and the determination of the manipulated variable. The purpose is to change the control characteristics. Weight Determination FIG. 37 shows the configuration of the weight determination unit provided in the ECU 20 of the above embodiment. The principle of this weight judgment is shown in FIG.
7. As shown in FIG. 38, the vertical acceleration signal is frequency analyzed, and the frequency component ω _n having the maximum power is detected from the analysis result. This ω _n is a quantity that reflects the total weight WGT of the vehicle.

In the above-mentioned three embodiments, two acceleration sensors are used. The weight determination unit shown in FIG. 37 outputs signals from these two sensors respectively to FFT (Fast Foul).
(ierTransfer) frequency analysis to detect ω _n1 and ω _n2 having the maximum power component in each power spectrum, and from the ω _n1 and ω _n2 using the determination table, corresponded to the current vehicle control amount Natural frequency ω _n and estimated maneuvering amount W
It is for detecting GT _n . This table is set in advance, and in the case of a right-hand drive vehicle, the output from the sensor on the left side is emphasized, and the output from the acceleration sensor on the rear part is also emphasized. Application to Control The natural frequency ω _n of the vehicle thus detected is used when the vertical velocity V _G is calculated from the vertical acceleration signal G. When calculating the vertical velocity V _G (step S4 in FIG. 9) from the vertical acceleration signal G,
As shown in FIG. 9, a bandpass filter and an integrator are used in combination. FIG. 40 shows the configuration of the integrator 60 shown in FIG. This supercharger, this integrator, has a plurality of bandpass filters, and each filter has a pass characteristic of ω ₁ to ω _m . Therefore, when the natural frequency ω _n is input to the integrator 60, a filter having a bandpass suitable for the value is selected, and its output becomes the vertical velocity V _G. By doing so, it is possible to calculate the vertical velocity V _G that reflects the actual vehicle weight at that time, that is, is accurate.

Output WGT from the weight judging unit 60_n
Is P that defines the upper limit of damping force_LMTUsed to calculate
It FIG. 41 shows a part of the front wheel control procedure of FIG. 9 (step
WGT from S22 to step S24)_nSo that you can add
It has been changed to. That is, in step S22, the vehicle speed V
_BDepending on P_LMTIs calculated, the stocker device 3 of FIG.
The correction coefficient k at 00 is k = κ · WGT_n/ WGT_B (18) is calculated. Where κ is a predetermined constant, and WGT_B
Is the vehicle weight when one driver of standard weight rides
is there. Therefore, WGT_n/ WGT_BIndicates the increase in weight
And changes as shown in FIG. That is, step S302
At P_LMT= P_LMT(V_B) .K ... (19) is calculated, P_LMTIs WGT_BMore weight than
In some cases P_LMT(V_B) Was increased from
WGT_BIf the weight is reduced than P_LMT(V _B)
Obtained as reduced. Thus, P
_LMTThe correction of
Etc., in the direction of changing the suspension characteristics to hard
It works as a positive.

Thus, it is possible to achieve both cost reduction and optimization of suspension characteristics. Modification Although the FFT is used in the above-described embodiment, the FFT is used in FIGS.
As shown in FIG. 4, a plurality of filters having different bandpass characteristics may be used, and the filter output having the maximum power may be selected.

Further, although the above-mentioned embodiment is an example of application to a so-called semi-active suspension device, the vehicle weight naturally has a great influence on the active suspension device, so that the above-mentioned ω _n and WG are
T can also be applied to an active suspension device. <Diversion to other systems of WGT> In automobiles, various controls are simultaneously performed in addition to suspension control. In some of these controls, the vehicle weight has a great influence on the control characteristics. For example, traction control, ABS control, rear wheel steering control are examples.

FIG. 45 shows an example in which the signal WGT is applied to the ABS controller. ABS Controller is EC
More accurate control can be performed using the signal WGT _n sent from U10. In this way, the G sensor required for suspension control can be used for vehicle weight determination, and the weight is also used for ABS control, so that the cost of the entire system can be reduced.

[0063]

As described above, since the total weight of the vehicle is reflected in the vertical acceleration, the natural frequency obtained by frequency analysis of the signal is the basis for estimating the vehicle weight. Therefore, if the estimated natural frequency is reflected in the control characteristic, it is possible to provide a vehicle suspension device that does not require a weight sensor and secures both low cost and suspension characteristic control in a high dimension. .

Since the natural frequency is a basis for estimating the total weight of the vehicle, the control characteristic is changed according to the total weight, for example, the suspension characteristic is corrected to the hard side according to the increase in weight. And so on. Furthermore, the weight data can be sent to another control system (for example, an ABS control system) to be used for the control system.

[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 a characteristic of a threshold V _G0 with respect to a vehicle speed V _B.

FIG. 11 is a graph showing the characteristic of the coefficient K ₁ with respect to the vehicle speed V _B.

FIG. 12 is a graph showing a characteristic of an upper limit value P _LMT of damping force with respect to a vehicle speed V _B.

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 P _R defined by a vehicle speed V _B in the first embodiment.

FIG. 15 is a graph showing a characteristic of a rear wheel target damping force P _RTR defined by the front wheel damping force P _OF 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 a characteristic of a front wheel target damping force P _F defined by a vehicle speed V _B in the second embodiment.

FIG. 21 is a graph showing the characteristics of the front wheel target damping force P _FTR defined by the rear wheel damping force P _OR 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 characteristics of the front wheel target damping force P _F with respect to the steering angle θ _H 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 V _B 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 G _A used in large amplitude input control
FIG. 6 is a graph showing the characteristics of the vehicle speed V _B.

FIG. 31: Coefficient G _B used in large amplitude input control
FIG. 6 is a graph showing the characteristics of the steering angle θ _H.

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 the characteristics of the coefficient G ₀ used for G-through control.

FIG. 35 is a graph showing the characteristics of the coefficient G ₁ used for G-through control.

FIG. 36 is a graph showing the characteristics of the coefficient G ₂ used for G-through control.

FIG. 37 is a diagram showing a configuration of a determination unit that estimates a vehicle weight and a natural frequency ω of the vehicle.

FIG. 38 is a view for explaining the principle of vehicle weight estimation.

FIG. 39 is a diagram showing the configuration of an integrator 60 that calculates the vertical velocity V _G using the estimated natural frequency ω.

FIG. 40 is a diagram showing a more detailed configuration of the integrator 60.

FIG. 41 is a flowchart showing a control procedure for using the estimated weight WGT to determine P _LMT .

FIG. 42 is a graph showing characteristics of WGT _n / WGT _B.

FIG. 43 is a diagram showing a configuration of a modified example of the determination unit.

FIG. 44 is a diagram illustrating the principle of a modified example of vehicle weight estimation.

FIG. 45 is a diagram illustrating a further application example of data WGT _n .

[Explanation of symbols]

1 _FL , 1 _FR ... high speed motor, 1 _RL , 1 _RR ... low speed motor,
1 _'FL, 1' _FR ... slow _{motor, 1 'RL, 1' RR} ... high-speed _{motors, 2 L, 2 R, 2} 'L, 2' R ... vertical G sensor, 60 ... integrator

Claims (7)

[Claims] 1. Inputting the detected vertical acceleration signal,
In a vehicle suspension device that feedback-controls a suspension characteristic of a vehicle in a direction in which the value of the acceleration signal decreases, a detection unit that detects a vertical acceleration signal and a suspension characteristic of the vehicle in a direction in which the value of the acceleration signal decreases. Control means for feedback control; analysis means for frequency-analyzing the detected vertical acceleration signal; determination means for determining the natural frequency of this vehicle from the analysis result; and the control means based on the determined natural frequency. And a means for changing the control characteristic of the vehicle suspension device. 2. The vehicle suspension device according to claim 1, wherein the control means includes a plurality of bandpass filter means having different filter characteristics, and the changing means includes the plurality of bandpass filter means based on the determined natural frequency. 2. The vehicle suspension device, wherein one of the bandpass filter means is selected, and the control means uses the output of the selected bandpass filter means for calculation of a feedback gain. 3. The vehicle suspension device according to claim 1, wherein the changing unit determines a total weight of the vehicle based on the determined natural frequency, and the control unit determines the total weight according to the determined total weight. A vehicle suspension device characterized by changing control characteristics. 4. The vehicle suspension device according to claim 3, wherein the changing means corrects the suspension characteristic to the hard side according to the weight increase when the determined total weight changes. A vehicle suspension device characterized by changing control characteristics. 5. The vehicle suspension device according to claim 3, wherein the data representing the determined weight is transmitted to another control system. 6. The vehicle suspension device according to claim 1, wherein the analyzing unit includes a Fourier transforming unit. 7. The vehicle suspension device according to claim 1, wherein the analyzing unit has a plurality of bandpass filter circuits.