JPH04100724A - Active stabilizer for vehicle - Google Patents

Abstract

PURPOSE:To conduct steering control preferentially without changing the whole roll and pitch rigidity by providing the first passage for connecting the pressure chambers to each other, the volume changes of which are the same to rolling, and the second passage for connecting the same similarly, the volume changes of which are the same to pitching. CONSTITUTION:The respective hydraulic cylinders 20FL-20RR are constructed so that the cylinder chambers U, L are communicated with each other in the close state by the first pipelines 26A, 26B and the second pipelines 27A, 27B. It is supposed that a vehicle transits to a left turning state. A controller 36 sets a motor turning angle command value ¦thetaM¦ corre sponding to the lateral acceleration ¦alphar¦ with reference to a map. When the electric motor 34 is turned counter-clockwise by an angle thetaM, a piston rod 30c is moved, so that the operating pressure of an upper chamber U of the cylinder 20RR, a lower chamber L of 20RL, a lower chamber of 20FR and an upper chamber U of 20FL is increased at the same time, and the operating pressure of the opposite chambers U, L is decreased at the same time. The axial force F due to such differential pressure P works on the cylinder rod 20c, and the magni tude thereof ¦ F¦ is substantially equal. In this arrangement, though the roll rigidity on the rear side is increased, it is decreased on the front side so that though the whole roll rigidity is not changed almost, the vehicle is changed to the oversteer side so as to heighten the turning round efficiency.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to an active stabilizer for a vehicle, and in particular, a double-acting hydraulic cylinder is individually inserted between four-wheel suspension links and a vehicle body. The present invention relates to an active stabilizer that controls the operating pressure of this fluid pressure cylinder.

[Conventional technology]

Conventionally, as a suspension having a function of suppressing rolling of a vehicle, the present applicant has proposed, for example,
910 (the name of the invention is "active suspension").

The suspension described in this publication has double-acting fluid pressure cylinders inserted between each wheel of the vehicle and the vehicle body, and the left and right fluid pressure cylinders have an upper pressure chamber on one side and a lower pressure chamber on the other side. The front and rear fluid pressure cylinders are connected to each other by cross piping, and the left wheel sides and right wheel sides are connected to each other by piping to form a piping system, and this piping system is connected to the steering torque of the power steering device. It is connected to a responsive control valve or a control valve linked thereto, and when turning, the fluid pressure cylinder on the outer race side is extended and the fluid pressure cylinder on the inner race side is contracted.

[Problem to be solved by the invention]

However, the main purpose of the above-mentioned conventional suspension is to control the total roll stiffness and suppress the roll angle by using cross piping between the left and right wheels, and although it is successful in that respect, the total Changing the roll stiffness would also change the steering characteristics, so it was not possible to control the steering characteristics alone. For example, when the sharing ratio of roll stiffness between the front and rear wheels set by coil springs is [10 (rear wheel): 5 (front wheel)], the roll stiffness is increased by "5" for the front and rear wheels by a total of "10". If this is done, the sharing ratio of roll stiffness during roll angle control will be 15 (rear wheels): 10 (front wheels), and the sharing ratio will change, so the steering characteristics will shift to the neutral steer side. etc., its characteristics will also change.

The present invention was made in view of this situation,
The problem to be solved is to actively control the front-rear distribution of roll stiffness or the left-right distribution of pitch stiffness without changing the overall roll stiffness or pitch stiffness, giving priority to control of steering characteristics and grip force. The purpose of the present invention is to provide a device that can perform the following tasks.

[Means for solving the problem] In order to solve the above problem, the invention described in claim (1) provides the following:
Double-acting hydraulic cylinders are swingably inserted between the suspension links of the four wheels and the vehicle body, and the pressure chambers of the hydraulic cylinders of the left and right wheels have the same volume change direction in response to rolling. A first flow path that interconnects the pressure chambers, and a first flow path that interconnects the pressure chambers that have the same volume change direction in response to pitching among the pressure chambers of the fluid pressure cylinders of the front and rear wheels on each of the left and right sides of the vehicle. An actuator section having a second flow path connected to the vehicle is provided, and an internal pressure control means is provided for changing the internal pressure of the actuator section according to the driving state.

[Operation] For example, suppose that you are aiming to control the steering characteristics when turning. Now, assume that each hydraulic cylinder has its cylinder tube connected to the vehicle body side and its piston rond connected to the wheel side, and the vehicle shifts to a left turn, causing rolling in which the right wheel side sinks. shall be. This results in
The internal pressure control means operates, for example, to increase the operating pressure in the upper pressure chamber of the hydraulic cylinder of the rear right wheel. At this time,
The pressure chambers of each of the four wheels' hydraulic cylinders cross at the left and right wheels, and are also connected at the front and rear wheels on the left and right sides of the vehicle, so that at the same time as the above-mentioned increase in operating pressure occurs, the pressure chambers of the rear left wheel's hydraulic cylinder also cross. This causes an increase in the operating pressure in the lower pressure chamber, the lower pressure chamber of the front right wheel fluid pressure cylinder, and the upper pressure chamber of the front left wheel fluid pressure cylinder. As a result, a pressure difference is generated between the side pressure chambers of each fluid pressure cylinder, and an axial force based on this pressure difference is generated in the cylinder rod. This axial force is
It points downward on the rear right wheel side, upward on the rear left wheel side, upward on the front right wheel side, and downward on the front left wheel side. Therefore, the roll rigidity on the front side decreases by an amount equal to the increase in the roll rigidity on the rear side, so the overall value is canceled out and the roll rigidity is maintained at a value determined by a coil spring or the like. However, as described above, the front-rear distribution of roll stiffness is higher on the rear wheel side, making it possible to actively control the steering characteristics toward oversteer.

The same applies when the vehicle turns to the right.

On the other hand, from a different perspective, the above explanation of the generation of axial force due to differential pressure is similarly applicable to the case where the left-right distribution of pitch rigidity is changed without changing the total pitch rigidity. This is advantageous in cases where only the grip forces of the left and right wheels of the vehicle are individually controlled.

〔Example〕

An embodiment of the present invention will be described below with reference to FIGS. 1 to 4 of the accompanying drawings.
This will be explained based on the diagram.

In FIG. 1, 2FL to 2RR indicate the front left wheel to the rear right wheel of the vehicle, 4 indicates a wheel support member, and 6 indicates the vehicle body.

One end of a suspension link 8 is swingably connected to the wheel support member 4, and the other end of the suspension link 8 is swingably connected to the vehicle body 6.

A vehicle suspension 9 is installed between each wheel support member 4 and the vehicle body 6, and this suspension 9 includes a shock absorber 10 and a coil spring 12 that are individually provided between each suspension link 8 and the vehicle body 6. , an active stabilizer 14 in which an actuator portion is provided between each suspension link 8 and the vehicle body 6.

Each of the shock absorbers 10 has a conventionally well-known configuration, and is configured to generate damping forces depending on the stroke speed on its compression side and expansion side, respectively.

The active stabilizer 14 also includes suspension links 8 for the front, left, and right wheels. ..., an actuator section 14A provided between the vehicle body 6 and the actuator section 14A, and the actuator section 14A.
and an internal pressure control means 14B for controlling the internal pressure of the internal pressure.

The actuator section 14A includes a hydraulic cylinder 2 as a fluid pressure cylinder disposed corresponding to each wheel 2FL to 2RR.
0FL~20RR and throttle valve 22A~ that generates damping force
22D, accumulators 24A to 24D for accumulating pressure,
1 for communication! Magnetic switching valve 25F.

25R, and each of these elements is interconnected by hydraulic piping.

Each of the hydraulic cylinders 20FL to 20RH has a cylinder tube 20a, and the inside of this cylinder tube 20a is separated into an upper cylinder chamber U (pressure chamber) and a lower cylinder chamber L (pressure chamber), and is slidable inside the tube. piston 2
0b, and a piston rond 20c fixed to the piston 20b and extending in the ring direction. Hydraulic cylinder 20 having such a structure
In each of FL to 20RR, the lower end of the piston rod 20c is swingably attached to the suspension link 8, and the upper end is left in a free state.
An end of the cylinder tube 20a in this example is swingably supported by the vehicle body 6, and thereby the hydraulic cylinder 2
0FL to 20RR are inserted into the front, rear, left and right sprung tops and unsprung bottoms, respectively.

Further, in each of the front wheel side and the rear wheel side, the upper cylinder chamber U of the left wheel side hydraulic cylinder 20FL (20RL) is connected to the right wheel side hydraulic cylinder 20PR via one first pipe 26A.
(20RR) is connected to the lower cylinder chamber of the left wheel side hydraulic cylinder 20FL (20RL), and the lower cylinder chamber of the left wheel side hydraulic cylinder 20FL (20RL) is connected to the right wheel side hydraulic cylinder 2 through the other first pipe 26B.
It is connected to the upper cylinder chamber U of 0PR (20RR), so that the cylinder chambers U and L are cross-connected to each other. The front and rear first pipes 26A, 26B are connected at an intermediate position via an electromagnetic switching valve 25F (25R). Each of the electromagnetic switching valves 25F and 25R is
Turning on the control signal S given to that solenoid.

It has a normally open (communication) structure that is in the closed and open positions depending on the off state.

Also, one first pipe 26A on the front wheel side is connected to a second pipe 27A.
The other first pipe 26B on the front wheel side is connected to the other first pipe 26A on the rear wheel side via a second pipe 27B. .

As a result, the front and rear hydraulic cylinders 20 on the left and right sides of the vehicle
The cylinder chambers U and L of FL, 20RL, 20PR, and 20RR communicate with each other in a cross state.

Here, the first pipes 26A and 26B correspond to the first flow path of the present invention, and the second pipes 27A and 27B correspond to the second flow path of the present invention.

A third pipe 28A is connected to the first pipes 26A and 26B of the front and rear wheels.
, 28B, of which the third piping 28A, 28B on the front wheel side is a throttle valve 22A.

22B to accumulators 24A and 24B, respectively. In addition, the third pipes 28A and 28B on the rear wheel side are connected to the throttle valve 2.
It reaches accumulators 24C and 24D via 2C and 22D, respectively.

On the other hand, the internal pressure control means 14B
A control cylinder 30 that applies an internal pressure of 4A, and a fourth pipe 32 connected to this control cylinder 30.
A, 32B, and an electric motor 34 for driving the control cylinder 30, as well as a controller 36 for control. Vehicle speed sensor 38. and a steering angle sensor 39.

Among these, the control cylinder 30 is a double cylinder like the aforementioned hydraulic cylinders 20FL to 20RR.

It is configured as a double acting type, and includes a cylinder tube 30a,
The inside of this cylinder tube 30a is divided into two cylinder chambers R1.
, R2 and a slidable piston 30b, and a piston rod 30c fixed to the piston 30b and extending in the ring direction. Among these, cylinder chamber R1,
R2 is connected to the third pipe on the rear wheel side via the fourth pipes 32A and 32B.
It communicates with piping 28A and 28B, respectively. Further, one end of the piston rod 30c is placed in a free state, and a rack 30d is formed at the other end. A pinion 34a of the electric motor 34 meshes with this hook 30d.

Further, the vehicle speed sensor 38 is, for example, a sensor that detects the rotation of the output shaft of a transmission, and outputs a pulse signal (2) to the controller 36 in accordance with the vehicle speed. The steering angle sensor 39 is a pulse detector installed in the steering system, and outputs a pulse signal θ corresponding to the steering direction and steering angle to the controller 36.

In this embodiment, the controller 36 includes a microcomputer, a motor drive circuit, a solenoid drive circuit, etc.
Detection signals (2) and (theta) from the vehicle speed sensor 38 and steering angle sensor 39 are input, the processing shown in FIG. A rotation angle sensor (not shown) is attached to the electric motor 34, and a motor rotation position signal θR from this sensor is supplied to the controller 36 to control the rotation position of the motor.

Next, the process shown in FIG. 2 executed by the controller 360 microcomputer will be explained. The process shown in the figure starts when the power is turned on.

To explain this, in step (2) in the figure, the microcomputer of the controller 36 reads the detection signals (2) and θ from the vehicle speed sensor 38 and steering angle sensor 39, stores the values as the vehicle speed and steering angle, and moves to step (3). . In this step ■, well-known calculations (for example, Japanese Patent Application Laid-Open No. 62-29316
7) is performed to estimate the lateral acceleration α7.

After this, the process moves to step (3), where the reference lateral acceleration α, which is near zero, is set in advance. For, IαYl>α. Determine whether or not. If this judgment results in rNOJ, it is assumed that control of wheel load movement is not necessary, and the process returns to step (2) via the process of step (2). In step ■, the control signals S, S
Both the front and rear electromagnetic switching valves 25 are turned off.
F, 25R is opened or kept open.

On the other hand, when rYEsj is determined in step (2), it is determined that such control is necessary, and the process moves to step (2). In this step (2), control signals S and S are both turned on, and the front and rear electromagnetic switching valves 25F and 25R are turned on.
Closed or keep it closed.

Next, the process moves to step (3), and with reference to the map corresponding to FIG. 3 stored in the memory in advance, the lateral acceleration 1α71
A motor rotation angle command value 1θ81 that is uniquely determined according to is calculated. The command value θ, 41 characteristic shown in FIG. 3 is designed to increase as the lateral acceleration α, 1 increases in order to suppress the roll angle.

Next, the process moves to step (2), and it is determined from the sign of the steering angle signal θ input in step (2) whether or not the steering wheel is turned to the right.

In the case of ``YES'' in this steering direction determination, steps ① to ② are performed. That is, the microcomputer outputs the motor drive signal i in the direction corresponding to the clockwise rotation of the motor (counterclockwise in FIG. 1) in step (3). Then, in step (2), the motor rotation position signal θR is input, and in step [phase], it is determined using the input signal θR whether the electric motor 34 has rotated in the right direction by the command value 08. If "NO", step ■.

The process of [phase] is repeated, and if rYEs, the motor rotation is stopped in step (2), and then the process returns to step (2). As a result, the electric motor 34 rotates in the right direction by the command value θ.

On the other hand, when determining "NO" in step ■, steps @~0. The process (2) is performed in the same manner as steps (2) to (2), thereby causing the electric motor 34 to rotate to the left by the command value θ.

Next, the overall operation of this embodiment will be explained.

Assume that the vehicle is traveling straight at a constant speed on a smooth road.

Since the controller 36 estimates the lateral acceleration αv', 0 with the start of the process shown in FIG. 2, the control signal is maintained at OFF and both the electromagnetic switching valves 25F and 25R are in an open (communicating) state. Further, in this straight-ahead state, since rotation of the electric motor 34 is not commanded, the piston 30b of the control cylinder 30 takes a neutral position, and the internal pressure of the actuator section 14A is not energized. As a result, predetermined suspension characteristics determined by suspension elements such as the coil spring 12 can be obtained without unnecessary load movement due to roll or pitch during non-control.

Assume that while the vehicle is traveling straight, both wheels bounce due to unevenness of the road surface. In this case as well, the estimated lateral acceleration αY is almost zero, so the switching valves 25F and 25R remain in communication, and the internal pressure of the actuator section 14A is not actively controlled and is controlled by each shock absorber 10. The damping force generated dampens the bounce. On the other hand, if the wheels 2PL and 2PR bounce due to passing through the convex portion and the strokes of the hydraulic cylinders 20FL and 20PR are about to be shortened, the upper cylinder chambers U of the cylinders 20FL and 20PR are simultaneously contracted, and the lower side Both cylinder chambers expand simultaneously. However, the compressed hydraulic oil in the cylinder chamber U flows through the pipes 26A, 26B and the switching valve 25F into the lower cylinder chamber of the same cylinder and the lower cylinder L of the opposite cylinder in almost equal amounts. Both chambers U.

Since the volume change of L is equal since both are Ronde shapes,
The amount of hydraulic oil overflowing from the cylinder chamber U is stored in the lower cylinder chamber with almost no difference, and the amount of oil in the second pipes 28A and 28B does not change. This is true even on the rear wheel side, and is also exactly the same when there is a rebound due to passing through a recess, only the vertical relationship is reversed.

Therefore, since the hydraulic oil does not pass through the throttle valves 22A to 22D during bounce and rebound, almost no damping force is generated, and a soft damping force characteristic is maintained, preventing deterioration of ride comfort when driving on uneven roads accompanied by bounce. are doing.

In addition, in such a straight-ahead state, one of the wheels 2P, for example,
Assume that a stroke variation occurs only in L due to riding over a protrusion or the like. In this case as well, there is no active internal pressure control of the actuator section 14A, but since the electromagnetic switching valve 2.5F is in communication state, the hydraulic oil overflowing from the cylinder chamber U is transferred to the switching valve 25F.
It flows into its own lower cylinder chamber through. In other words,
Piping 28A, 28 due to the bypass action of switching valve 25F
No oil amount change occurs in B, and no damping force occurs. Therefore, even in the case of such one-wheel bounce, riding comfort is hardly impaired.

Now assume that the vehicle has transitioned from the straight-ahead state described above to a turning state on a good road. For example, suppose this turning is a left turning as shown in Fig. 4, and when viewed from the rear of the vehicle, the vehicle body on the right wheels 2FR and 2RR sinks, and the left wheels 2FL and 2RL try to lift up. . During this turn, the controller 36 estimates and calculates the lateral acceleration α7 with a sign corresponding to the left turn from the steering angle θ and the vehicle speed ■, and sets the motor rotation angle command value 1θ, which is uniquely determined corresponding to the lateral acceleration 1αv1, to 1. Set by map reference.

Then, the controller 36 determines left-turn steering from the sign of the steering angle signal θ, and rotates the electric motor 34 to the left (clockwise in FIG. 1 in the present example) by an angle of 09 minutes. Forced by this rotation, the piston rod 30c moves toward the left end in FIG. 1 by an amount corresponding to the rotation angle θH.

As a result, one cylinder chamber R1 of the control cylinder 30 is compressed and the pressure in the acid chamber R1 increases, while the other cylinder chamber R2 is expanded and the pressure in the acid chamber R2 decreases.

As a result, the upper cylinder chamber U of the rear right hydraulic cylinder 20RR, the lower cylinder chamber of the rear left hydraulic cylinder 20RL, the lower cylinder chamber of the front right hydraulic cylinder 20PR, and the upper side of the front left hydraulic cylinder 20FL. The working pressure in the cylinder chamber U increases at the same time (shaded area in FIG. 4), and the working pressures in the cylinder chambers U and L on the opposite side simultaneously decrease. As a result, a pressure difference ΔP is generated between the upper and lower cylinder chambers U and L in each of the hydraulic cylinders 20FL to 20RR, and an axial force ΔF due to this pressure difference ΔP acts on the cylinder chew 20c.

This axial force ΔF is applied to the rear right side as shown by the arrow in Fig. 4.
It acts downward (toward the road surface) on the front left side, and upward (toward the vehicle body) on the rear left side and front right side, and
The magnitude 1ΔF1 is almost the same. Therefore, the reaction force acting on the vehicle body is in the opposite direction to the arrow in the figure.

Now, since the front right wheel and the rear right wheels 2FR and 2RR are on the outer wheel side, the moment due to the rear side axial force ΔF will be in a direction that resists rolling, and the rear side roll rigidity will increase.
On the contrary, the direction on the front side corresponds to rolling, and the roll rigidity on the front side decreases. but,
Since each axial force 1ΔF1 is almost the same, the increase in roll rigidity on the rear side and the decrease in roll rigidity on the front side cancel each other out, and the overall roll rigidity hardly changes, and each shock absorber 10. ..., 10 generates a damping force and suppresses posture fluctuation, and the roll angle is 4-wheel coil spring 12. ..., 12, etc., is controlled to a desired value by the roll rigidity.

However, because the roll stiffness decreases on the front side and increases on the rear side, the front-rear distribution ratio of roll stiffness increases on the rear side, and as a result, the steering characteristics change to more oversteer side. , the turning ability is improved.

Furthermore, in the above-described control, by changing the magnitude of the load f on the control cylinder 30, that is, the motor rotation angle θ, even if the roll rigidity is increased or decreased, the amount of increase or decrease can be finely adjusted. . Specifically, in this embodiment, the rate of change of the command value characteristic curve shown in FIG. 3, which is the map data, may be changed.

The above-mentioned control is the same when turning left.

Therefore, for example, the initial stage of the turn and the time of the convergence of the turn are determined based on the rate of change of the estimated value of the lateral acceleration α7, and when determining the initial stage of the turn, internal pressure control is performed to increase the rear roll rigidity, and when determining the convergence of the turn, the internal pressure is controlled in the direction of increasing the rear roll rigidity. It is preferable to control the internal pressure in a direction that increases the roll rigidity on the front side. As a result, while the roll angle is held substantially constant, it is possible to change the steering characteristics to oversteer at the beginning of a turn to improve turning performance, and to understeer at the end of a turn to ensure turning stability.

Furthermore, in this embodiment, the internal pressures at the front and rear of the vehicle are simultaneously controlled by a single internal pressure control means 14B, so the manufacturing cost is reduced due to the simplified configuration compared to the case where the internal pressures at the front and rear of the vehicle are controlled by separate control means. At the same time, the weight can also be reduced.

Furthermore, when viewed from a different perspective, the internal pressure control of this embodiment also controls the left and right distribution of pitch stiffness at the same time. For this reason, when braking, for example, if the internal pressure control for a left turn as described above is performed (see Figure 4), the pitch rigidity of the left wheel increases when the front side of the vehicle is about to sink, and the right wheel increases by that amount. The side pitch stiffness decreases. As a result, without changing the overall pitch rigidity determined by the suspension elements such as the coil spring 12,
The grip force on the left wheel side is lowered, and the grip force on the right wheel side is increased. Moreover, the left-right distribution of pitch rigidity can be adjusted by the force f applied to the control cylinder 30, that is, the motor rotation angle θ, so that when the vehicle is turning, accelerating, decelerating, and spinning (during acceleration) or tuck-in, It is effective to adopt this pitch stiffness distribution when the vehicle exhibits a tendency of (during deceleration).

Note that the configuration of the actuator section of the present invention is not limited to that described in the above-described embodiment, and may be as shown in FIG. 5, for example. In the same figure, front left wheel 2P
L, rear left wheel 2RL side double acting cedar hydraulic cylinder 20FL, 2
In ORL, the lower end of the cylinder tube 7'20a is swingably connected to the suspension link 8, and the upper end of the piston rod 20c is swingably connected to the vehicle body 8. Also, since the double-acting cedar hydraulic cylinders 20PR and 20RR on the front right wheel 2FR and rear right wheel 2RR sides are inserted in the same manner as in the above-mentioned embodiment, the directions of the hydraulic cylinders are opposite for the left and right wheels. . The first pipes 26A and 26B connect the lower cylinder chambers and the upper cylinder chambers U of the left and right cylinders 20FL, 20PR, 20RL, and 20RR, respectively. In addition, the second pipe 2
7A is the first pipe 26A on the front wheel side and the first pipe 2 on the rear wheel side.
6B, and the other second pipe 27B connects the first pipe 26B on the front wheel side and the first pipe 26A on the rear wheel side. The rest of the configuration is the same as the embodiment described above.

As a result, the piping structure appears to be parallel both on the left and right sides and in the front and back, but since the insertion directions of the left and right hydraulic cylinders are reversed, it is possible to obtain the same effect as in the above-mentioned embodiment. can.

Further, the internal pressure control means in the present invention may use a lateral acceleration sensor that directly detects lateral acceleration.

Furthermore, the actuator section in the present invention includes the front side and rear side switching valves 25F in the above-described embodiment,
It is also possible to have a structure in which 25R is omitted.

Furthermore, the working fluid in the present invention is not limited to the one that uses hydraulic oil as described above, but may be an apparatus that uses, for example, incompressible gas as the working fluid.

〔Effect of the invention〕

As described above, the claimed invention includes a double-acting hydraulic cylinder that is swingably inserted between the suspension links of the four wheels and the vehicle body and has the same specifications; A first flow path that interconnects pressure chambers in which the volume changes in the pressure chambers of the fluid pressure cylinder have the same direction of increase and decrease in response to rolling, and a volume in the pressure chambers of the fluid pressure cylinders of the front and rear wheels that corresponds to pitching. The actuator section is provided with a second flow path that interconnects the pressure chambers in which the direction of increase and decrease is the same, and the internal pressure control means is provided that changes the internal pressure of the actuator section according to the running state. , without changing the total roll stiffness.
It is possible to actively control the front-rear distribution of roll stiffness to change steering characteristics, and to change the grip force distribution by actively controlling the left-right distribution of pitch stiffness without changing the total pitch stiffness. Along with the front.

Since the internal pressure on the rear side can be controlled simultaneously by a single internal pressure control means, the weight and cost can be reduced compared to the case where internal pressure control means are provided individually.

[Brief explanation of the drawing]

1 to 4 are diagrams showing one embodiment of the present invention, in which FIG. 1 is a schematic configuration diagram of the entire device, FIG. 2 is a schematic flow chart showing an example of processing in the controller, and FIG. 3 is a diagram showing an example of the present invention. 4 is a lateral acceleration characteristic diagram of the motor rotation angle command value stored as map data, and FIG. 4 is an explanatory diagram showing the direction in which axial force is generated by applying a load to the control cylinder. FIG. 5 is a schematic diagram of the entire apparatus showing another embodiment. Main symbols in the diagram are 6... Vehicle body, 8... Suspension link, 14... Active stabilizer, 14A.
...actuator section, 14B...internal pressure control means, 2
0FL~20RR...Hydraulic cylinder, 26A, 26B
...First pipe, 27A, 27B...Second pipe, 30
... Control cylinder, 32A, 32B... Fourth pipe, 34... Single motion motor, 36... Controller, 38... Vehicle speed sensor, 39... Steering angle sensor,
It is. Rate rain forward ○ side # love! L gourd Y

Claims (1)

[Claims] (1) Double-acting hydraulic cylinders are swingably inserted between the suspension links of the four wheels and the vehicle body, and changes in volume due to rolling among the pressure chambers of the hydraulic cylinders of the left and right wheels. A first flow path that interconnects pressure chambers having the same direction; and a pressure chamber whose volume change in response to pitching increases or decreases in the same direction among the pressure chambers of the fluid pressure cylinders of the front and rear wheels on each of the left and right sides of the vehicle. an actuator section having a second flow path that interconnects the actuator sections;
An active stabilizer for a vehicle, characterized in that it is provided with an internal pressure control means that changes the internal pressure of the actuator section according to the driving condition.