JP3046817B1 - Vehicle suspension method and apparatus - Google Patents

Abstract

An object of the present invention is to improve the ride comfort of a vehicle by damping a suspension of the vehicle in each mode, and to improve the maneuverability and stability of the vehicle by controlling the posture of the vehicle well. SOLUTION: A load generating actuator 18 composed of a hydraulic cylinder is disposed between each wheel 14 and a vehicle body 16, and a tire vertical displacement and a vertical displacement speed are detected for each wheel. When the load generating actuator 18 corresponding to each wheel 14 is controlled so that the vehicle body balances left, right, up and down based on the tire vertical displacement speed, the vehicle can be subjected to vibration isolation over all of pitching, rolling and bouncing. Ride comfort and maneuverability can be improved.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle suspension and, more particularly, to an improvement in a vehicle suspension method and apparatus for supporting a vehicle body so as to stabilize a posture of the vehicle when the vehicle is changing on a road surface or turning. It is.

[0002]

2. Description of the Related Art Prior art vehicle suspension apparatus 1
00 is a coil spring 102 arranged between each wheel 14 of the vehicle 12 and a vehicle body (not shown) as shown in FIG.
And a shock absorber 104. Vehicle 1
2 vibrates up and down during running in response to changes in the road surface, the knuckle 106 moves up and down via the spindle in the radial direction and the axle direction of the tire 14A without causing relative movement with respect to the tire. The upper and lower upper link 108 and the lower link 110 are respectively arranged in the front-rear direction of the vehicle body around the upper link pin 112 and the lower link pin 114, that is, the vertical direction between the tire 14A and the vehicle body with the longitudinal direction of the vehicle 12 as the rotation axis. A relative displacement occurs.

When the vehicle suspension device 100 is of a parallel double wishbone type, as shown in FIG. 22, an upper link length Lu and a lower link length L1,
That is, when the distances in the vehicle width direction between the upper link pin Pu and the lower link pin P1 and the upper and lower ball joints BJu and BJl of the knuckle are equal, the line L1 connecting the upper and lower link pins and the upper and lower ball joints of the knuckle are connected. The connecting line L2 is displaced up and down in parallel.

As described above, the upper link 108 or the lower link 110 is connected to the respective link pins 112, 1
When the rotational displacement is rotated up and down about the rotation axis 14, the rotational displacement is transmitted to the mounting points of the coil spring 102 and the shock absorber 104 attached to the vehicle body, and expands and contracts the coil spring 102 and the shock absorber 104.
A transmission load is generated on the vehicle body side according to the expansion and contraction speed of the shock absorber 104, and the tire 14A vibrates up and down while absorbing kinetic energy with respect to the vehicle body side.

When the suspension device 100 is of a four-wheel double wish board type and has no stabilizer, the vertical movement of one of the four tires 14A and the corresponding load generated by the coil spring 102 and the shock absorber 104 corresponding thereto. Are one-to-one and independent of the vertical movement of the other three tires.

Lower links 110 at both ends of the vehicle body in the vehicle width direction
When the stabilizer 116 is installed over the tire, a spring load is generated for the relative displacement of the left and right tires 14A so as to suppress the vertical displacement, but no load is generated for the relative speed. . It should be noted that the stabilizer 116 acts so that the vertical displacement of each tire is affected only by the relative displacement of the left and right tires.

[0007] However, the conventional suspension device has various problems. First, when a vehicle travels on a good road, the sprung part (body side) vibrates with pitching, rolling and bouncing as natural vibration mode displacements, and the movement on the spring is the superposition of these natural vibration mode displacements. Expressed as a sum. As described above, while the natural vibration mode of the vehicle has three modes, the vibration characteristics of the spring and the damper (shock absorber that generates damping force) constituting the conventional suspension device are different from those of the front and rear suspensions. It only has two forms of each shock absorber. For this reason, even if the vibration of pitching is changed by adjusting the damping coefficient of the front and rear shock absorbers, the other vibrations also change, so that the three vibrations of pitching, rolling and bouncing cannot be adjusted independently. Therefore, there were drawbacks in that the ride quality was poor and the anti-vibration speed was low.

When the suspension device is a parallel double wishbone type, the right and left tires are displaced up and down due to bouncing and pitching, and when the upper arm and the lower arm are rotationally displaced, the tire tread changes and the tire wear is accelerated. Was causing it.

In the parallel double wishbone suspension system, the shock absorber is arranged in the vehicle height direction, and the upper link length, the lower link length, and the stabilizer are arranged in the vehicle width direction. The vehicle cannot be reduced in the vehicle height direction and the vehicle width direction, and the vehicle space cannot be effectively used.

Further, the prior art suspension device is
As described below, there are disadvantages in controlling the attitude of the vehicle. (1) At the time of turning of the vehicle, as the vehicle speed is increased and the turning radius is reduced, the center of gravity of the vehicle deviates outside the center track of the tread, so that the vehicle easily falls over and the maneuverability of the vehicle decreases. . (2) When the vehicle is suddenly braked, the vehicle leans forward, the rear wheels rise off the road surface, the braking force is reduced, and a sharp turn occurs. (3) When the vehicle suddenly starts, the rear portion of the vehicle sinks, and the riding comfort is reduced.

[0011]

One of the problems to be solved by the present invention is to improve the ride comfort of the vehicle by means of vibration isolation by mode, and to control the vehicle attitude satisfactorily by controlling the vehicle attitude. It is an object of the present invention to provide a vehicle suspension method and device capable of improving performance.

Another object of the present invention is to prevent a change in tread, which tends to be caused by a vertical displacement of a tire, to prevent the tire from being worn, and to improve the maneuverability of a vehicle. It is an object of the present invention to provide a vehicle suspension method and apparatus that can be used.

Still another object of the present invention is to provide a mechanism for transmitting a generated load to a tire in a small space, thereby reducing the size of the vehicle in the vehicle width direction and the vehicle height direction. It is an object of the present invention to provide a vehicle suspension method and apparatus capable of performing the above.

[0014]

SUMMARY OF THE INVENTION A first object of the present invention is to provide a cylinder chamber having a pressure between a wheel and a vehicle body.
A load generating actuator composed of a fluid cylinder that generates a load on the piston according to the difference is arranged, and on the other hand, the tire vertical displacement and the vertical displacement speed are detected for each wheel, and the tire vertical displacement and the tire vertical displacement speed of all wheels are detected. The present invention is to provide a vehicle suspension method characterized by controlling load generating actuators corresponding to respective wheels so that the vehicle body is balanced in front, rear, left and right and up and down based on the control parameters.

According to a second aspect of the present invention, there is provided a vehicle suspension method according to the first aspect,
The center-of-gravity acceleration of the vehicle is further detected, and a load generating actuator corresponding to each wheel is used as a control parameter by adding the center-of-gravity acceleration to the tire vertical displacement and the tire vertical speed of all the wheels, and each wheel is determined based on these three control parameters. The present invention provides a suspension method for a vehicle, characterized by controlling a load generating actuator corresponding to (1).

According to a third aspect of the present invention, there is provided a method for controlling a pressure difference between a cylinder chamber and each of a wheel and a vehicle body.
A load generating actuator composed of a fluid cylinder that generates a load on a piston, a tire vertical displacement sensor that detects a vertical displacement of a tire for each wheel, and a tire vertical displacement speed sensor that detects a vertical displacement speed of a tire for each wheel And control for controlling the load generating actuators of the respective wheels such that the vehicle body is balanced in the front-rear, left-right, up-down directions based on the control parameters using the detection values of the tire vertical displacement sensors and the tire vertical displacement speed sensors of all the wheels as control parameters. And a suspension system for a vehicle, comprising:

According to a fourth aspect of the present invention, there is provided a vehicle suspension device according to the third aspect,
The control system further includes a center-of-gravity acceleration detection unit that detects a center-of-gravity acceleration of the vehicle, and the control system adds a detection value of the center-of-gravity acceleration detection unit to a detection value of a tire vertical displacement sensor and a tire vertical speed sensor of all wheels as a control parameter. It is an object of the present invention to provide a vehicle suspension device that controls a load generating actuator corresponding to each wheel based on these three control parameters.

According to a fifth aspect of the present invention, there is provided a vehicle suspension device according to the third or fourth aspect, wherein a load generating actuator corresponding to each wheel includes:
A suspension device for a vehicle, comprising: a torque transmission mechanism that is arranged to be shifted in a longitudinal direction of the vehicle with respect to a tire position of a corresponding wheel and that transmits torque between a load generating actuator and the wheel. Is to provide.

According to a sixth aspect of the present invention, there is provided a vehicle suspension apparatus according to any one of the third to fifth aspects, wherein the control system is configured to control the vertical displacement of the tire and the vertical displacement speed of the vehicle body. It is another object of the present invention to provide a suspension device for a vehicle, wherein the suspension spring constant and the damping constant are changed by superimposing the displacement of the center of gravity, the speed of the center of gravity, and the acceleration of the center of gravity on the side or by applying a manual operation.

According to a seventh aspect of the present invention, there is provided a vehicle suspension apparatus according to any one of the third to sixth aspects, wherein the load generating actuator generates a load determined based on each control parameter. A plurality of fluid cylinder groups, each fluid cylinder group includes a fluid cylinder for each vibration mode, and a plurality of fluid cylinder groups and fluid cylinders in each fluid cylinder group are tandemly arranged. An object of the present invention is to provide a suspension device.

Thus, between each wheel of the vehicle and the vehicle body ,
A load generating actuator composed of a fluid cylinder that generates a load on the piston according to the pressure difference of the cylinder chamber is arranged, and detects the tire vertical displacement and the vertical displacement speed for each wheel, and detects the tire vertical When the load generating actuators corresponding to the wheels are controlled such that the vehicle body is balanced in the front-rear, left-right, up-down directions based on the control parameters, the displacement and the tire vertical displacement speed as control parameters,
Vibrations on the vehicle body side can be prevented for each natural vibration mode, so that the vehicle can be damped over all of pitching, rolling and bouncing, and the riding comfort and maneuverability of the vehicle can be significantly improved.

Further, when the vehicle body vibrates up and down, the distance between the left and right wheels does not change, and the tire does not wear due to the change in the tire tread. Is preferable.

Further, when each load generating actuator is arranged to be shifted in the longitudinal direction of the vehicle with respect to the tire position of the corresponding wheel, and is connected to the wheel by a torque transmitting mechanism, the space occupied by the suspension becomes The vehicle can be reduced not only in the vehicle width direction but also in the vehicle height direction, so that the vehicle can be downsized or the space in the vehicle can be effectively used.

In particular, the suspension spring constant and the damping constant can be changed by superimposing the displacement of the center of gravity, the speed of the center of gravity, and the acceleration of the center of gravity of the vehicle body without changing the tire in addition to the vertical displacement and the vertical displacement speed of the tire or by adding a manual operation. When it is changed, the ride comfort of the vehicle can be further improved, and at the same time, the exercise performance such as maneuverability can be improved.

Also, the vehicle attitude can be freely adjusted by reducing the amount of change of the zero reference point of the vertical displacement of the tire from the vertical displacement of each tire by manual operation or automatic operation according to the acceleration of the center of gravity. This is to prevent the limit moment of the vehicle from decreasing at the time of turning, thereby improving the maneuverability and fallability of the vehicle, preventing the vehicle from turning forward during braking, and reducing the braking distance. The sinking of the rear portion of the vehicle at the time of starting can be prevented, and the riding comfort can be improved.

In particular, a large vehicle such as a commercial vehicle having a large center of gravity shift has a large change in vibration and posture in various modes, but changes the suspension spring constant and the damping constant according to the center of gravity shift at the time of starting and braking. In addition, the anti-vibration effect and attitude control are remarkably large, and therefore, it is very effective especially for a large vehicle.

The load generating actuator provided on each wheel is composed of a plurality of fluid cylinder groups that receive a generated load determined based on each control parameter, and each fluid cylinder group includes a fluid cylinder for each vibration mode. When the plurality of fluid cylinder groups and the fluid cylinders in each fluid cylinder group are arranged in tandem, the control parameters and the generated load for each vibration mode can be easily obtained, and the responsiveness is improved. A high-quality control system having reactivity can be easily designed.

[0028]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of a vehicle suspension apparatus 10 according to the present invention will be described in detail with reference to the drawings. FIGS. 1 and 2 schematically show a suspension apparatus 10 according to the present invention. This suspension device 10
Are load generating actuators 18A to 18D (hereinafter collectively referred to as “hydraulic cylinders”) arranged between respective wheels 14A to 14D of the vehicle 12 (hereinafter, generally denoted by reference numeral 14) and the vehicle body 16 respectively. A tire vertical displacement sensor 20 (refer to FIG. 2) for detecting a vertical displacement d of the tire for each wheel 14;
The detection values of the tire vertical displacement speed sensor 22 (see FIG. 2) for detecting the vertical displacement speed v of the tire every four wheels, and the detection values of the tire vertical displacement sensors 20 and the tire vertical displacement speed sensors 22 of all the wheels 14 are used as control parameters. And a control system 26 for controlling the load generating actuators 18 of the respective wheels so that the vehicle body 16 is balanced in the front-rear, left-right and up-down directions based on the control parameters.

The control system 26 can also control the load generating actuator 18 based on the detected value of the center-of-gravity acceleration g _{0 in addition to} the tire vertical displacement d and the tire vertical displacement speed v. As shown in FIG. 3, the center-of-gravity acceleration g ₀ is obtained by calculation from four acceleration sensors 24 arranged in the front-rear and up-down directions within the symmetrical vertical plane VF of the vehicle 12, as shown in FIG. And will be described later.

A specific embodiment of the suspension device 10 of the present invention will be described in detail with reference to FIG. 4 and the following figures. FIGS. 4 to 6 show a non-drive system (a front wheel of a rear wheel drive vehicle) to which the present invention is applied. Side) one front wheel 14A of the suspension device 10
10A corresponding to FIG.
Is a spindle (axle) 3 that supports one front wheel 14A.
A torque transmission mechanism 28 for transmitting a torque between the zero and the load generating actuator 18 is provided. The other front wheel 14
The device part corresponding to B is constructed in exactly the same way as the device part 10A.

As shown in FIG. 5, the spindle 30 has a substantially C-shaped spindle hanger 31 integrally therewith.
The spindle hanger 31 has a swing pin 3
2, 32 ', and these rocking pins 32, 32'
Keys 38, 38 'provided on the inner surfaces of the bosses 36, 36' between the boss 36 of the upper support stay 34 and the boss 36 'of the lower support stay 34' make it non-rotatable but slidable. The spindle hanger 31 is swingably supported by being engaged with the supported upper and lower bush-shaped sliding members 40 and 40 ′. Accordingly, the spindle 30 swings about the axis of the upper and lower sliding members 40 and 40 ', and the front wheel 14A (or 14
B) can be directed in the turning direction of the vehicle. The upper and lower support stays 34, 34 'are shown in FIG.
As shown in the figure, the arm is composed of forked arms, and the flanges 34a and 34'a at the ends of the arms are screwed to a vehicle body (not shown) and attached to the vehicle body. FIG.
7 to FIG. 7, reference numerals 40a and 40'a denote key grooves provided on the outer surfaces of the upper and lower sliding members 40 and 40 ', respectively, in the vertical direction, and into which the keys 38 and 38' are engaged.

The torque transmission mechanism 28 includes a spindle-side movable member 42 having upper and lower horizontal mounting pieces 43 and 43 ′ attached to opposed ends of upper and lower support stays 34 and 34 ′, and a load generating actuator 14. It comprises an actuator-side movable member 44 having an attached horizontal mounting piece 44a, and interlocking means 46 for interlocking these movable members 42 and 44 with each other.

The horizontal mounting piece 4 of the movable member 42 on the spindle side
Bosses 45, 45 'supported by the sliding member 40 by holding the bush-shaped sliding member 40, respectively.
Having. These bosses 45, 45 'consist of two boss halves 45A, 45B, 45'A, 45'B, and these boss halves 45A, 45B have bush-like sliding members 40, 45B therein. The flanges 45a, 45b are screwed to each other so as to hold the sliding member 40 '.
It is supported by 40 '.

Accordingly, the spindle side movable member 42 is restrained from being displaced in the horizontal direction with respect to the vehicle body 16 by the upper and lower support stays 34 and 34 'for supporting the upper and lower sliding members 40 and 40'. The actuator-side movable member 44 includes rollers 48A and 48B provided in front and rear directions.
And 48′A, 48′B restrict the displacement in the horizontal direction with respect to the vehicle body 16.

In the illustrated embodiment, the interlocking means 46 includes racks 50, 52 provided on the outer surfaces of the movable members 42 and 44.
And an idler gear 54,
And a torque transmission shaft 62 having pinion gears 58 and 60 meshing with each other via a gear 56. The tire vertical displacement sensor 20 and the tire vertical displacement speed sensor 22 are provided on the torque transmission shaft 62.

The device for suspending the other front wheel 14B of the rear-wheel drive vehicle also has the same configuration as the device 10A of FIGS. 3 to 7, and a description thereof will be omitted. In FIG. 4, reference numeral 63 denotes a bearing attachment position that restricts parallel movement of the triangle vertex in the vertical and horizontal directions, but allows rotational movement around the triangle vertex.

Next, a suspension device 10CD of a drive system (the rear wheels 14C and 14D on the drive side of the rear wheel drive vehicle).
This suspension device 10CD is substantially the same as the device portion 10A, except that a drive shaft 64 for driving the rear wheels 14C and 14D is supported through the spindle hanger 31. The drive shaft 64 supports the respective wheels 14C and 14D via a differential gear device (not shown), so that the device portion 10A
And different. The same portions as those of the device portion 10A are denoted by the same reference numerals.

In particular, as can be seen from FIG. 9, the movable member 42 is vertically supported at a position shifted laterally from the drive shaft 64 so that the spindle-side movable member 42 does not interfere with the drive shaft 64. It is supported by stays 34, 34 '. 8 and 9, the drive shaft 64 crosses the torque transmission shaft 62 in a three-dimensional manner so that the drive shaft 64 does not interfere with the torque transmission shaft 62. In FIG. 9, reference numeral 65 denotes a bearing attachment position at which movement in the left-right direction is allowed by rollers below the triangle.

When the vehicle 12 having the non-drive system and the drive system suspension devices 10A and 10CD is subjected to vibration and the tires of the respective wheels 14 are vertically displaced, the up-and-down displacement and the displacement speed become the spindle 30 and the spindle hanger 31.
And movable member 4 via upper and lower support stays 34, 34 '.
2, and further transmitted to the torque transmission shaft 62 via the rack 50, the idler gear 54, and the pinion gear 58. Therefore, the vertical displacement d and the vertical displacement speed v of the tire of each wheel 14 are detected by the vertical displacement sensor 20 and the vertical displacement speed sensor 22 provided on the torque transmission shaft 62. Is input to

The vertical displacement sensor 20 is composed of, for example, a linear encoder capable of detecting a subtle change in the vertical position of the torque transmission shaft 62 using a change in light, magnetism or capacitance. Vertical displacement speed sensor 2
2, for example, periodic irregularities, optical slits or NS
It consists of a scale velocimeter that detects alternating magnetic changes mechanically, optically or magnetically or a velocimeter using the Doppler effect.

The acceleration sensor 24 shown in FIG. 3 for obtaining the center-of-gravity acceleration Ga is composed of, for example, a piezoelectric accelerometer for detecting acceleration by a piezoelectric effect under the force of a weight.
The acceleration g ₀ at the position G of the center of gravity is represented by four acceleration sensors 24
Detection values g ₁ to g ₄ from the height h ₁ to h ₄ from the road surface of the acceleration sensor 24, the height h from the road surface of the center of gravity G, the distance in the longitudinal direction from the gravity center position G of the acceleration sensor 24 The following equation (1) is obtained by the control system 26 from W ₁ , W ₂ (the distance W ₁ between the upper and lower acceleration sensors is the same, and the distance W ₂ between the upper and lower acceleration sensors is also the same).
Is calculated from

[0042]

(Equation 1)

As shown in FIG. 12, the control system 26 includes signals (detected values) from the sensors 20, 22, and 24 and setting signals (target values of the spring constant and the damping coefficient) set by the driver manually. The microcomputer 66 includes a microcomputer 66 for generating a control signal for controlling each load generating actuator 18 so as to balance the vehicle body 16 in the front, rear, left, right, up and down directions. In FIG. 12, reference numeral 68 denotes an A / D signal when the outputs of the sensors 20 to 26 are analog signals.
An A / D converter for converting, and reference numeral 68 'denotes a D / A converter for D / A converting a control signal when the hydraulic control circuit of the load generating actuator 18 is an analog circuit. Reference numerals 70I and 70O denote input ports and output ports of the microcomputer 66, respectively.

Load generating actuators 18A to 18D
13 receives the hydraulic pressure from a hydraulic pump 74 driven by an electric motor 72 and generates each load from a microcomputer 66, as shown in FIGS. Actuator 18
It is activated by hydraulic circuits 78A to 78D (see FIG. 1) including control valves 76 which receive control signals for controlling A to 18D, respectively. The control valve 76 is provided for each load generating actuator.
The corresponding load generating actuators 18A to 18D are controlled based on the command from the microcomputer 66, but the commands from the microcomputer 66 and the load generating actuators 18A to 18D are controlled.
The load is set so as to be proportional to the generated load of 8D.

The microcomputer 66 includes the sensor 2
By reading the sensor signals from 0, 22 and 24 and the setting signal set by manual operation, the load generated by the load generating actuator 18 which is the output device is calculated, and the load is applied to the piston by the pressure difference between the left and right hydraulic cylinder chambers. Generate. The calculation of the generated load is performed by the following equation (4).
7) to (50).

Next, the specific operation of the suspension device 10 of the present invention will be described below together with the relationship between the load generated by each load generating actuator 18, the vertical displacement of each tire, and the vertical displacement speed.

First, the load generated by the load generating actuator 18 and the vertical displacement and the vertical displacement speed of each tire are expressed by the following natural equations (2) and (3).
become that way. The symbols used in the equations (2) and (3) are as follows. (Zj) Displacement vector (Mij) Mass tensor (Ki
j) Stiffness tensor (Pj) Load vector (ω) Natural frequency (Iij)
Unit tensor (subscript i, j, k, l, m) natural number

[0048]

(Equation 2)

[0049]

(Equation 3)

If Pi = KijZj, the eigenmode of the eigenfrequency is obtained as in the following equations (4) and (5).

[0051]

(Equation 4)

[0052]

(Equation 5)

On the other hand, the arbitrary load is expressed as a superposition sum of the eigenmode loads similarly to the eigenmode displacement at an arbitrary displacement. Next, a forced load motion equation on the vehicle body side (spring-up) of the ideal vehicle is expressed by the following equations (6) to (8). The symbols used in equations (6) to (8) are as follows. (Z) bouncing amount ([theta] x) roll angle ([theta] y) pitch angle (F) forcing the external force (M) Forced moment (M _P b) vehicle body mass (C _R) centroid rolling modal damping factor (K _R) centroid rolling mode spring constant (C _P) centroid pitching mode damping coefficient (K _P) centroid pitching mode spring constant (ippr) (= ipro + M P pb · h 2) the vehicle body side roll moment of inertia (iPrO) centroid vehicle body roll moment of inertia around ( Ipp) (= Ippo + Mpb · h ² ) Vehicle-side pitch inertia moment (Ippo) Vehicle-side pitch inertia moment around center of gravity (Subscript p) Pitching (Subscript _B ) Bouncing (Subscript _R ) Rolling (Subscript _L ) Left (Subscript _R ) ) Right side (subscript f) Front wheel
(Subscript r) rear wheel (subscript x) forward coordinate of vehicle (subscript y) leftward coordinate of vehicle (subscript z) upward coordinate of vehicle

[0054]

(Equation 6)

[0055]

(Equation 7)

[0056]

(Equation 8)

When the equations (6) to (8) are rewritten from the vertical displacement d detected by the vertical displacement sensor 20 and the vertical displacement speed v detected by the vertical displacement speed sensor 22 for each tire, the bouncing is represented by the following equation: 9) and (10), equations for rolling are given by equations (11) and (12), and pitching are given by equations (13) and (14). The symbols used in equations (9) to (14) are:
It is as follows. (H) Tire tread (Wf) Distance from center of gravity position to front wheel axle (Wr) Distance from center of gravity position to rear wheel axle (WB = Wf + Wr) Wheelbase (axle distance)

[0058]

(Equation 9)

[0059]

(Equation 10)

[0060]

[Equation 11]

[0061]

(Equation 12)

[0062]

(Equation 13)

[0063]

[Equation 14]

From the equations (6), (9) and (10), the inertia force in the vehicle-body-side center-of-gravity bouncing mode is represented by the following equation (15). Further, from the vehicle-body-side skeleton model shown in FIG. The inertia of the center of gravity bouncing mode is
When each tire reaction force is represented by the load generated by the actuator, the expression (16) is obtained. The symbols used in equations (15) and (16) are as follows. (C _B) barycenter bouncing mode damping coefficient (K _B) barycenter bouncing mode spring constant (Pf _R) right front side of the actuator generated load (Pf _L) the left front wheel side of the actuator generated load (Pr _R) of the right rear wheel side actuator Generated load (Pr _L ) Load generated by the actuator on the left rear wheel side

[0065]

(Equation 15)

[0066]

(Equation 16)

Similarly, from the equations (7) and (11) and (12), the inertia force in the vehicle body side rolling mode is
Equation (17) is obtained, and the inertial force in the vehicle body-side rolling mode from the vehicle body-side skeleton model in FIG. 15 is expressed by equation (18).

[0068]

[Equation 17]

[0069]

(Equation 18)

Similarly, from the equations (8), (13) and (14), the inertia force in the vehicle body side pitching mode is as shown in equation (19), and the equation (19) is obtained from the vehicle body side skeleton model of FIG. 20).

[0071]

[Equation 19]

[0072]

(Equation 20)

When the center-of-gravity load vector is represented by a mode load vector, the equation (21) is obtained.

[0074]

(Equation 21)

The first to third terms on the right side of equation (21)
Are the bouncing mode, rolling mode and
And load in pitching mode. Equations (16), (1
8) and (20) and the load of the bouncing mode alone and
Actuator generated load component P _Bf_R, P_Bf_L, P_Br_R,
P_Br_LSymmetry, ie, P_Bf_R= P_Bf_L, P_Br_R= P_B
r_LFrom these, the load components are given by the following equations (22) (2)
3)

[0076]

(Equation 22)

[0077]

(Equation 23)

Similarly, equations (16), (18) and (2)
0) a rolling mode alone of the load and the actuator generating load component _{P R f R, P R f} L, P R r R, P R r before and after the load balance of _L, that _{P R f L = (Wr /} Wf) P _R r _{L,
P R f R = (Wr /} Wf) P R r from _R of load components, the following equation (24) (25).

[0079]

(Equation 24)

[0080]

(Equation 25)

Similarly, equations (16), (18) and (2)
0) and the load and the actuator generating load component of the pitching mode alone _{P P f R, P P f} L, P P r R, P P r L DOO symmetry, i.e. _{P P f R = P P f} L, P P From r _R = P _P r _L , these load components are represented by the following equations (26) and (27).

[0082]

(Equation 26)

[0083]

[Equation 27]

Therefore, the combined value of the load generated by the actuator corresponding to each wheel is given by the following equations (28) to (31).

[0085]

[Equation 28]

[0086]

(Equation 29)

[0087]

[Equation 30]

[0088]

(Equation 31)

Equations (15), (17), and (19) are replaced by Equation (2).
A _{B1 to} a _B3 are substituted into Equations (2) to (31) to obtain Equation (3).
2) to (34), a _{R1 to} a _R3 are set as in equations (35) to (37), and a _P is set as in equation (38).

[0090]

(Equation 32)

[0091]

[Equation 33]

[0092]

(Equation 34)

[0093]

(Equation 35)

[0094]

[Equation 36]

[0095]

(37)

[0096]

(38)

Therefore, the calculated values of the load generated by the actuator corresponding to each wheel are given by the following equations (39) to (42).

[0098]

[Equation 39]

[0099]

(Equation 40)

[0100]

[Equation 41]

[0101]

(Equation 42)

To compare this with the suspension characteristics of a prior art combination of a shock absorber and a spring, the load generated for each wheel according to this prior art is calculated by the following equations (43) to (46). It is on the street. The symbols used in equations (43) to (46) are as follows. (Cf) Front wheel damper characteristic value of the current vehicle (Cr) Rear wheel damper characteristic value of the current vehicle (Kf) Front wheel spring constant of the current vehicle (Ks) Front wheel stabilizer spring constant of the current vehicle

[0103]

[Equation 43]

[0104]

[Equation 44]

[0105]

[Equation 45]

[0106]

[Equation 46]

Equations (39) to (42) are converted to vibration parameters C _B , C _R , and C which are the damping coefficient and the spring constant of each mode.
_P, K _B, K _R, and rearranging each K _P, the following equation (47) to (50).

[0108]

[Equation 47]

[0109]

[Equation 48]

[0110]

[Equation 49]

[0111]

[Equation 50]

The actual load generated on each wheel is calculated by the above equation (47).
(48) The load obtained in (49) and (50) is superimposed on the load shared by each vehicle when the vehicle is at rest, and is synthesized. The microcomputer 66 in the control system 26 generates this load by generating each load. Control to act on the actuator. In both suspension systems of the non-drive system and the drive system, the input load of the vertical displacement of the tire and the load generated by the actuator are superimposed or subtracted, and the vehicle body is controlled to have a stable posture.

In the vehicle equipped with the suspension device of the present invention, it is assumed that the vertical displacement dfL has occurred on the left front wheel, but the vertical displacement has not occurred on the other wheels, and the damping coefficient and the spring in the rolling mode and the pitching mode have been changed. Assuming that the constant is zero, the above equations (47) to (50) in which the actuator generated load of each wheel is obtained,
These are represented by the following equations (51) to (54), respectively.

[0114]

(Equation 51)

[0115]

(Equation 52)

[0116]

(Equation 53)

[0117]

(Equation 54)

Therefore, Pf _L = Pf _R , Pr _L = Pr _R (Pf _L + Pf _R ) Wf = (Pr _L + Pr _R ) Wr, and it is understood that no rotational moment is generated around the center of gravity.

For example, in a conventional independent suspension type device having no stabilizer, a vertical load P1 due to the deflection of a suspension spring based on the vertical displacement of one wheel.
On the other hand, the vertical load P2 due to the deflection of the suspension spring at the other adjacent wheel breaks the balance of P2 <P1, and therefore the vertical load caused by the deflection of the suspension spring with respect to the arm length based on the offset amount from the center of gravity. A load P is generated, and a rotational moment is generated around the center of gravity using the amount of offset from the position of the center of gravity as the length of the arm. According to the present invention, as described above, it is advantageous that all the wheels are controlled so that this rotational moment is canceled.

In Formulas (47) to (50), while the vehicle is running, the driver manually sets the damping coefficients and the spring coefficients C _B , C _R , C _P , K _B , K _R , and K _P in each mode. By varying, the posture on the vehicle body side can be controlled to be more stable. These also, as already mentioned, besides the vertical displacement d and the vertical displacement speed v of each tire,
When the vehicle turns, the sudden braking and sudden acceleration or the like, it is possible to control the load generating actuator of each wheel gravity acceleration g ₀ as control parameters.

In this way, it is possible to prevent a decrease in the limit moment of the vehicle at the time of a sharp turn, thereby improving the overturning property, and preventing the rear wheels from rising during the sudden braking to improve the stability and the braking performance. In addition, it is possible to prevent the ride comfort from deteriorating due to the sinking of the rear portion at the time of sudden start.

As already mentioned, each load-generating actuator 18 can be shifted back and forth in the longitudinal direction of the vehicle with respect to the tire position of the corresponding wheel 14, for example to be arranged longitudinally inward with respect to the tire position. (See FIGS. 1, 4 and 8), the space occupied by the suspension device 10 can be reduced not only in the vehicle width direction but also in the vehicle height direction.
The space inside can be used effectively. This is because, for example, in the prior art device of the independent suspension type, the vertical position of the hood of the vehicle is limited to secure the upper fixed position of the shock absorber, and the vehicle width required to secure the link arm length is Compared to having a space in the direction,
It turns out to be advantageous.

Furthermore, since the tires move up and down in parallel with the vertical direction of the vehicle body when rolling the vehicle, the distance between the wheels remains unchanged even when rolling, and the tire tread coincides with the distance between the wheels during bouncing and pitching of the vehicle. Therefore, the tire is not worn due to the change of the tire tread. This is because the link arm is used in the conventional independent suspension system of the prior art, so that the tire tread changes due to the rotation of the link arm during bouncing and pitching, and it is significantly superior to the tendency that the tire tends to wear. I understand.

The load generating actuator 18 of each wheel is
The vertical displacement and the vertical displacement speed at the four tire (wheel) positions are controlled by inputting the vertical displacement and the vertical displacement speed. Therefore, for example, by setting the pitching and rolling spring constants that cause the vehicle body rotation input to zero by manual operation, the flatness is controlled. Riding on bad roads can be achieved while maintaining a comfortable ride.

It is noted that if the vibration condition of the vehicle is an external force such as lateral acceleration or air resistance acting on the center of gravity of the vehicle, this becomes a rotation input of the vehicle, so that vibration isolation by mode is important. Should. When the longitudinal and lateral accelerations of the vehicle body act, the forcing moments of equations (7) and (8) are
M _R ≠ 0 and M _P ≠ 0, and the center of gravity rolling mode damping coefficient C _R and the center of gravity rolling mode spring constant K _R are C _R
= K _R = 0, and the center-of-gravity pitching mode damping coefficient _CP and the center-of-gravity pitching mode spring constant K _P cannot be C _P = K _P = 0.

Further, when the vehicle posture is adjusted so that the lateral acceleration G such as centrifugal force and the gravitational acceleration g are balanced when the vehicle turns, the acceleration to the human body can be reduced and the riding comfort can be improved. This is, as shown in FIG.
Attitude control can be achieved by manual operation or automatic operation by gravitational acceleration g so that θ = G / g.

Next, the movable member 44 supporting the load-generating actuator 18 with respect to the actuator-generated load.
Is considered with reference to FIG. As shown in FIG. 16A, when the actuator generated load P is applied to the center of the movable member 44 (the center in the thickness direction),
As shown by the dotted line in FIG.
Assuming that the angle between the contact surface between the rack 52 on the movable member 44 and the idler gear 56 and the horizontal plane is α, FIG.
As shown in (1), a load component of P1 = Psinα and P2 = Pcosα acts between the rack 52 and the idler gear 56. The load component P1 is a load component perpendicular to the contact surface between the rack 52 and the idler gear 56, and the load component P2 is a component perpendicular to P1.
Is a component that produces a slip on the contact surface between the motor and the idler gear 56.

Therefore, the movable member 4 is moved from the idler gear 56 to the movable member 4.
17, a load of Psin α · cos α acts on the movable member 44, and as shown by the dotted line in FIG. 17A, the four-point supported rollers 48A, 48B and 48′A,
While being supported by 8′B, the side surface of the roller is deformed so as to be convex on the side opposite to the side having the idler gear 56.

On the other hand, as shown in FIG. 17, when α is set to 0 ° and the load P generated by the actuator is shifted from the center of the movable member 44 to the side opposite to the idler gear 56 by a distance e, the movable member 44 becomes It is deformed in a direction opposite to that of FIG.

Therefore, α ≠ 0, the load position of the generated load P is shifted from the center of the movable member 44 by e, and
When supported at four points by 8A, 48B and 48'A, 48'B, the movable member 44 becomes a deformation of FIG.
8, and the bending of the movable member 44 in the vehicle width direction is suppressed. Therefore, the tooth tips of the rack 52 and the idler gear 56 do not come into contact with each other, the sliding resistance of the movable member 44 is reduced, and an increase in weight due to excessive constraint of four or more rollers can be suppressed.

Each load generating actuator 18 is shown in FIG.
As shown in FIG.
It is preferable that (1) to (6) are arranged in tandem. In this way, for example, the generated load determined based on the vertical displacement speed v is applied to the first hydraulic cylinder group including the three hydraulic cylinders 19 (1) to 19 (3), and the other three The generated load obtained based on the vertical displacement d can be applied to the second hydraulic cylinder group including the hydraulic cylinders 19 (4) to 19 (6). Further, the three hydraulic cylinders 19 (1) to 19 (3) or 19 (4) to 19 (6) of each group apply the generated load determined according to the three vibration modes of bouncing, rolling and pitching, respectively. receive. still,
19, reference numerals 76 (1) to 76 (6) denote control valves for controlling the respective hydraulic cylinders 19 (1) to 19 (6), and reference numeral 77 denotes these control valves 76.
This is a pipeline connected to (1) to 76 (6).

Thus, the load generated on each load generating actuator 18 is the sum of the loads Pr _{R1 to} Pr _R6 shared by the hydraulic cylinders 19 (1) to 19 (6). It is represented by equation (55). Further, the respective shared loads Pr _{R1 to} Pr _R6 are represented by equations (56) to (61), respectively.

[0133]

[Equation 55]

[0134]

[Equation 56]

[0135]

[Equation 57]

[0136]

[Equation 58]

[0137]

[Equation 59]

[0138]

[Equation 60]

[0139]

[Equation 61]

In equations (56) to (61),
The values of a _B2 , a _B3 , a _R2 , a _R3 , and a _p are determined by the values of Wf and Wr used to determine them when the vehicle is stationary.
Generated load P of each actuator when _R ≠ 0 and K _P ≠ 0
It is obtained by calculation using r _R , Pr _L , Pf _R , and Pf _L.

Similarly, Pr _L , Pf _R and Pf _L are obtained, and the generated loads on the front wheel side and the rear wheel side are obtained from the following equations (62) and (62). Note that Por in the last term in equations (56) to (61)
_R is the design load when the vehicle is empty, and is Pr _L , Pf _R and Pf _L
Similar design load Po _L r in, Pof _R and Pof _L is used.

[0142]

(Equation 62)

[0143]

[Equation 63]

When the load generating actuators of all the wheels are controlled in consideration of the generated load for every vibration mode parameter, the vehicle can always run stably. And maintainability, and the reliability and serviceability of the vehicle can be improved. Further, it is understood that the generated load instructed by the computer 66 in the control system 26 can be dispersed to improve the responsiveness of the load and control can be performed quickly.

Note that d and v are based on the tire rebound, so that d> 0 and v> 0 at the time of rebound.
Therefore, Pr _R1 and Pr _R4 decrease, and therefore, the overall Pr _R also decreases.

The most important characteristic of the present invention is that a load generating actuator composed of a fluid cylinder (hydraulic cylinder) is used instead of the combination of the spring and the shock absorber of the prior art suspension, and the tires detected on all wheels are used. Is to control load generating actuators corresponding to all wheels so as to maintain a stable posture of the entire vehicle from these control parameters using the vertical displacement d and the displacement velocity v as control parameters.

As described above, the suspension using the load generated in each load generating actuator as a control item is based on the principle of balance between the mass point and the load of the degree of freedom vibration system.
That is, the difference between the inertia force of the mass point and the external force F balances in the opposite direction to the superimposed amount of the spring restoring force and the center-of-gravity damping force. The spring restoring force is proportional to the vertical displacement of the mass point, and the gravity center damping force is proportional to the vertical displacement speed of the mass point. Load generating actuator 18
Generates a load component to be balanced with the difference between the inertial force of the mass point and the external force F.

As described above, in order to generate a load that balances the difference between the inertial force of the mass point and the external force, it is difficult to combine the spring restoring force proportional to the vertical displacement and the damping force proportional to the vertical displacement speed. Essentially different from technology suspension. For example, like an active suspension generally known as an FI suspension,
Although a hydraulic cylinder is used instead of the spring and the damper,
Since this active suspension does not use such a principle of balance, it is essentially different from the suspension method of the present invention.

Further, in the present invention, if a spring is used in addition to the hydraulic cylinder between the vehicle body and the wheels, a flat posture cannot be maintained when traveling on a rough road at a low speed. Can not achieve the swing. Therefore, the present invention is essentially different from the method of controlling the attitude of the vehicle body with respect to the support (cab) by using both a hydraulic cylinder and a spring as described in JP-A-6-278648. is there.

In the above-described embodiment, the case where the control parameters of the generated load are the vertical displacement and the vertical displacement speed of the tire has been described in detail. However, as described above, these control parameters can be obtained by equation (1). can be further added gravity acceleration g ₀ as control parameters. This corresponds to, for example, the six hydraulic cylinders 19 (1) in FIG.
In each of the vibration modes, control is performed by controlling three load hydraulic actuators by cascading three hydraulic cylinders 19 each having the center of gravity acceleration as a control parameter.

Further, the spring constant and the damping constant of the suspension can be continuously changed by the driver's manual operation or the displacement of the center of gravity of the vehicle, the speed of the center of gravity and the acceleration of the center of gravity. Useful for improving athletic performance such as ride comfort and steering feeling. The basic requirement for attitude control of the vehicle body is that the suspension load static load increases when the suspension stroke is unchanged, and that the suspension load static load does not change when the suspension stroke is variable. Athletic performance and riding comfort corresponding to the change in posture can be improved.

In the above-described embodiment, the load is generated by supplying the hydraulic pressure to the hydraulic cylinder, which is the load generating actuator, via the control valve. However, as shown in FIG. It comprises a hydraulic cylinder 80 having hydraulic chambers 82, 82 'closed on both sides of a piston rod 80a.
82 ′ may be communicated by a conduit 86 including an on-off valve 84, and a pinion 92 connected to an output shaft of an electric motor 90 may be engaged with a rack 88 integrated with the piston rod 80 a.

[0153]

According to the present invention, as described above , the piston is provided between each wheel of the vehicle and the vehicle body by the pressure difference in the cylinder chamber.
A load generating actuator composed of a fluid cylinder that generates a load on each ton is disposed, a tire vertical displacement and a vertical displacement speed are detected for each wheel, and based on the tire vertical displacement and the tire vertical displacement speed of all wheels. By controlling the load generating actuators corresponding to each wheel so that the vehicle body balances left, right, up and down, vibrations on the vehicle body side can be prevented according to the natural vibration mode, so that the vehicle is damped over all pitching, rolling and bouncing. Therefore, the riding comfort and maneuverability of the vehicle can be significantly improved.

Further, when the vehicle body on the suspension vibrates up and down, the distance between the left and right wheels does not change, and the tire does not wear due to the change in the tire tread. It is preferable because the vibration of the above is suppressed.

Furthermore, if each load generating actuator is arranged so as to be displaced not in the vehicle width direction of the vehicle but in the longitudinal direction with respect to the tire position of the corresponding wheel, the space occupied by the suspension becomes smaller in the vehicle width direction. Of course, the size can be reduced in the vehicle height direction, and the size of the vehicle can be reduced, or the space in the vehicle can be effectively used.

In particular, in addition to the vertical displacement and vertical displacement speed of the tire, the suspension spring constant and the damping constant can be reduced by superimposing the displacement of the center of gravity, the velocity of the center of gravity, and the acceleration of the center of gravity of the vehicle without changing the tire or by adding a manual operation. When it is changed, the ride comfort of the vehicle can be further improved, and at the same time, the exercise performance such as maneuverability can be improved.

Further, by reducing the amount of change of the zero reference point of the vertical displacement of the tire from the vertical displacement of each tire by manual operation or automatic operation according to the acceleration of the center of gravity, the posture of the vehicle can be freely adjusted. This is to prevent the limit moment of the vehicle from decreasing when turning, thereby improving the maneuverability and fallability of the vehicle, preventing the vehicle from turning in front of the vehicle during braking, and reducing the braking distance. The sinking of the rear portion of the vehicle at the time of starting can be prevented, and the riding comfort can be improved.

In particular, a large vehicle such as a commercial vehicle having a large center of gravity shift greatly changes vibrations and postures in various modes, but changes the suspension spring constant and the damping constant according to the center of gravity shift at the time of starting and braking. In addition, the anti-vibration effect and attitude control are remarkably large, and therefore, it is very effective especially for a large vehicle.

The load generating actuator provided on each wheel is composed of a plurality of fluid cylinder groups receiving a generated load determined based on each control parameter. Each fluid cylinder group includes a fluid cylinder for each vibration mode. When a plurality of fluid cylinder groups and the fluid cylinders in each fluid cylinder group are arranged in tandem, the control parameters and the generated load for each vibration mode can be easily obtained, and the responsiveness is improved. A high-quality control system having reactivity can be easily designed.

[Brief description of the drawings]

FIG. 1 is a perspective view of a skeleton of a vehicle suspension device according to the present invention.

FIG. 2 is a system diagram showing a relationship between a load generating actuator constituting one device part of a suspension device and a corresponding wheel.

FIG. 3 is a perspective view for explaining mounting positions of four acceleration sensors necessary for detecting a center of gravity acceleration.

FIG. 4 is a perspective view of a suspension device portion of a non-drive system.

FIG. 5 is an enlarged cross-sectional view taken along line 5-5 ′ of the suspension device of FIG. 4;

FIG. 6 is an enlarged top view of the suspension device of FIG. 4 as viewed from an arrow 6;

FIG. 7 is an exploded perspective view of an upper support stay of the suspension device of FIG. 4;

FIG. 8 is a perspective view of a suspension device portion of a drive system.

9 is an enlarged cross-sectional view taken along line 9-9 'of the suspension device shown in FIG.

FIG. 10 is an enlarged top view of the suspension device shown in FIG.

FIG. 11 is an exploded perspective view of the upper support stay of the suspension device of FIG. 8;

FIG. 12 is a system diagram showing a relationship between control means, input means (sensor), and output means (load generation actuator) used in the present invention.

FIG. 13 is a block diagram showing one load generating actuator and its driving means.

FIG. 14 is a system diagram of a specific example of the driving means of FIG.

FIG. 15 is an explanatory diagram showing a generated load (suspension reaction force) of a corresponding actuator of each wheel on a skeleton model of a vehicle.

FIG. 16 is an explanatory diagram for balancing a heavy acceleration g with a lateral acceleration G acting when the vehicle turns.

17A and 17B illustrate one deformation state of the load generating actuator, FIG. 17A illustrates the principle of the deformation, and FIG. 17B illustrates the force relationship of a meshing portion between the rack and the idler gear. FIG.

FIG. 18 is a diagram illustrating the principle of another deformation state of the load generating actuator.

FIG. 19 is a system diagram illustrating a specific example of one load generating actuator.

FIG. 20 is another explanatory view of the load generating actuator used in the present invention.

FIG. 21 is a perspective view of one example of a prior art suspension device.

FIG. 22 is a diagram illustrating displacement of upper and lower links of a conventional parallel wishbone suspension device.

[Explanation of symbols]

 REFERENCE SIGNS LIST 10 suspension device 10A device portion corresponding to one front wheel 10CD device portion corresponding to rear wheel 12 vehicle 14 wheels 14A one front wheel 14B other front wheel 14C one rear wheel 14D the other rear wheel 16 body 18 load generating actuator 18A one A load generating actuator 18B corresponding to the other front wheel 18B A load generating actuator 18C corresponding to the other front wheel 18D A load generating actuator 18D corresponding to the other rear wheel 19 (1) Hydraulic cylinder 19 (2) Hydraulic cylinder 19 (3) Hydraulic cylinder 19 (4) Hydraulic cylinder 19 (5) Hydraulic cylinder 19 (6) Hydraulic cylinder 20 Tire vertical displacement sensor 22 Tire vertical displacement speed sensor 24 Acceleration sensor 26 Control system 28 Torque transmission mechanism 30 Speed Dollar (axle) 31 Spindle hanger 32 Swing pin 32 'Swing pin 34 Upper support stay 34a Flange 34' Lower support stay 34'a Flange 36 Boss 36 'Boss 38 Key 38' Key 40 Bush-shaped sliding member 40a Key groove 40 'Bush-shaped sliding member 40'a Key groove 42 Spindle-side movable member 43 Horizontal mounting piece 43' Horizontal mounting piece 44 Actuator-side movable member 44a Horizontal mounting piece 45 Boss 45A Half boss 45B Half boss 45 'Boss 45'A Boss half 45'B Boss half 46 Interlocking means 48A Roller 48B Roller 48'A Roller 48'B Roller 50 Rack of movable member 42 Rack 52 of movable member 44 Idler gear 56 Idler gear 58 Pinion gear 60 Pinion gear 62 Torque Reaching shaft 63 bearing mounting position 64 drive shaft 65 bearing mounting position 66 microcomputer 68 A / D converter 68 ′ D / A converter 70I input port 70O output port 72 electric motor 74 hydraulic pump 76 control valve 76 (1) control valve 76 (2) Control valve 76 (3) Control valve 76 (4) Control valve 76 (5) Control valve 76 (6) Control valve 78A Hydraulic circuit 78B Hydraulic circuit 78C Hydraulic circuit 78D Hydraulic circuit 80 Hydraulic cylinder 80a Piston rod 82 Hydraulic Chamber 82 ′ Hydraulic chamber 84 On-off valve 86 Pipe line 88 Rack 90 Electric motor 92 Pinion 100 Conventional suspension device 102 Coil spring 104 Shock absorber 106 Knuckle 108 Upper link 110 Lower link 112 Upper link pin 114 Lower link pin Lu Up Link length Lu Li lower link length Pu upper link pin Pi lower link pin BJu upper ball joint BJl lower ball joint

Claims (7)

(57) [Claims] A cylinder chamber is provided between each wheel of a vehicle and a vehicle body.
A load generating actuator composed of a fluid cylinder that generates a load on the piston by the pressure difference of
On the other hand, the tire vertical displacement and the vertical displacement speed are detected for each wheel, and the vehicle body is balanced in the front-rear, left-right and up-down directions based on the control parameters using the tire vertical displacement and the tire vertical displacement speed of all the wheels as control parameters. Controlling a load generating actuator corresponding to each wheel. 2. The vehicle suspension method according to claim 1, further comprising detecting a center-of-gravity acceleration of the vehicle, and setting a load generating actuator corresponding to each wheel to a tire vertical displacement and a tire vertical displacement of all the wheels. A suspension method for a vehicle, wherein a control parameter is obtained by adding the acceleration of the center of gravity to a speed, and a load generating actuator corresponding to each wheel is controlled based on the three control parameters. 3. A load is generated on a piston by a pressure difference in a cylinder chamber disposed between each wheel of the vehicle and the vehicle body.
A load generating actuator comprising a fluid cylinder,
Tire vertical displacement sensor that detects the vertical displacement of the tire for each wheel, tire vertical displacement speed sensor that detects the vertical displacement speed of the tire for each wheel, tire vertical displacement sensor and tire vertical displacement speed sensor for all wheels And a control system for controlling the load generating actuators of the respective wheels such that the vehicle body is balanced in the front-rear, left-right, and up-down directions based on the control parameters using the detected values as control parameters. apparatus. 4. The vehicle suspension device according to claim 3, further comprising a center-of-gravity acceleration detecting means for detecting a center-of-gravity acceleration of the vehicle, wherein the control system includes a tire vertical displacement sensor for all the wheels. A control value obtained by adding a detection value of the center-of-gravity acceleration detection means to a detection value of a tire vertical speed sensor, and controlling a load generating actuator corresponding to each wheel based on the three control parameters. Suspension device. 5. The vehicle suspension device according to claim 3, wherein the load generating actuators corresponding to the respective wheels are arranged to be shifted in the longitudinal direction of the vehicle with respect to the tire positions of the corresponding wheels. And a torque transmission mechanism for transmitting torque between the load generating actuator and the wheels. 6. The vehicle suspension device according to claim 3, wherein the control system includes a vertical displacement and a vertical displacement speed of the tire, a displacement of a center of gravity of the vehicle body, a center of gravity speed, and a center of gravity. A suspension device for a vehicle, wherein a suspension spring constant and a damping constant are changed by superimposing acceleration or by applying a manual operation. 7. The vehicle suspension device according to claim 3, wherein the load generating actuator includes a plurality of fluid cylinder groups receiving a generated load determined based on each control parameter, A vehicle suspension apparatus, wherein each fluid cylinder group includes a fluid cylinder for each vibration mode, and the plurality of fluid cylinder groups and fluid cylinders in each fluid cylinder group are arranged in tandem.