JPH04135907A - Wheel load shift control device - Google Patents

Abstract

PURPOSE:To enhance the control effect for the steering characteristic of a vehicle, by arranging a hydraulic stabilizer or the like on the rear wheel side if the steering characteristic of the vehicle has an oversteering tendency upon rolling, but on the front wheel side if it has an understeering tendency. CONSTITUTION:If the steering characteristic of a vehicle given by a suspension device including a coil spring 12 as a main component and the like varies in an oversteering (understeering) direction upon rolling, a hydraulic stabilizer 14 is attached to the front (rear) wheel side. Accordingly, an equi-steering characteristic curve which descends rightward can be drawn on two-dimensional coordinates between total roll rigidity and roll rigidity. When the hydraulic stabilizer 14 is actuated, the steering characteristic ascends rightward, passing through a characteristic point given by the suspension device, and accordingly, the control width of the steering characteristic becomes wider, including a larger number of equi-steering characteristic curves, than that obtained by an arrangement in which the stabilizer is attached on the rear (front) wheel side so that the steering characteristic descends rightward.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention is a wheel load transfer control device for a stabilizer, etc., which controls the amount of load transfer between wheels that occurs due to rolling of a vehicle body. Regarding.

[Conventional technology]

BACKGROUND ART Conventionally, as a device for controlling the amount of movement of a wheel load, for example, the device described in Japanese Utility Model Application Laid-open No. 60-765 '06 (name of the device is "hydraulic stabilizer") is known.

This conventional device has single-route and double-acting hydraulic cylinders installed in the vertical direction between the left and right suspension arms and the vehicle body, and between the left and right hydraulic cylinders, one upper cylinder chamber and the other The lower cylinder chamber is communicated with the lower cylinder chamber through hydraulic piping in an intersecting state, and orifices are inserted in the middle of each hydraulic piping, and the hydraulic piping portion between the upper cylinder chamber and the orifice of each hydraulic cylinder is
A spring mechanism that elastically biases the hydraulic oil is communicated.

As a result, when the vehicle body rolls, a pressure difference in the opposite direction is generated between the upper and lower cylinder chambers of the left and right hydraulic cylinders, and this creates a force that resists rolling, and the orifice A damping force is generated based on the aperture effect.

[Problem to be solved by the invention]

However, the above-mentioned conventional device is configured to passively control the amount of load transfer between the left and right wheels simply in accordance with the lateral acceleration acting on the vehicle. Since it was installed in In other words, if a wheel load transfer control device is installed on only one of the front and rear wheels, and the roll stiffness on one side is changed, the total roll stiffness will also change, so the roll when turning with the same lateral acceleration The angle also changes. This roll angle change is caused by a general vehicle (
(Roll steer and camber angle changes occur due to rolling), the vehicle's steering characteristics also change due to roll steer and camber angle changes. If this change in steering characteristics goes against the control direction of the steering characteristics that is intended to be controlled based on the roll stiffness distribution, the total control width of the steering characteristics will be narrowed, and the effect of controlling the steering characteristics will not be fully exerted. I was in a situation where I wasn't.

The present invention has been made in view of the problems of the conventional devices, and the problem to be solved is to install a wheel load movement control device in consideration of the direction of change in steering characteristics due to changes in roll angle, The objective is to always maximize the control effect of the steering characteristics by the control device.

[Means to solve the problem]

In order to solve the above problem, the claimed invention includes an actuator whose roll rigidity can be changed between the left and right wheels at either the front or rear of the vehicle and the vehicle body, and by changing the roll rigidity of this actuator. In a wheel load transfer control device that controls the amount of load movement, if the vehicle steering characteristics due to the suspension device provided between the four wheels and the vehicle body change in the oversteer direction during rolling, one of the left and right wheels is replaced by the front wheel. If the steering characteristic of the vehicle caused by the suspension device changes in the understeer direction during rolling, one of the left and right wheels is set as the rear wheel.

[Operation] When the vehicle steering characteristics caused by the suspension device change toward understeer during rolling, a wheel load transfer control device is attached to the rear wheel side.

Now, when the roll stiffness distribution increases on the rear side due to the operation of the wheel load transfer control device, the steering characteristics change in the direction of oversteer. On the other hand, when the total roll rigidity is increased (that is, the roll angle is small), the steering characteristics of the suspension device change in the direction of oversteer.

Therefore, if we express the above relationship in two-dimensional coordinates using the total roll stiffness as the horizontal axis (the up direction is to the right of the horizontal axis) and the vertical axis of the roll stiffness distribution (the increase in the rear side is to be the upward direction of the vertical axis), the above relationship will be described later. As shown in FIG. 4, it is possible to draw an equal steering characteristic line that slopes downward to the right on the coordinates. and,
When the wheel load transfer control device installed on the rear wheel side is activated,
The steering characteristic is upward-sloping to the right after passing through the characteristic points of the suspension device, and the steering characteristic includes more equal steering characteristic lines than the downward-sloping characteristic when the wheel load transfer control device is assumed to be installed on the front wheel side. The control range becomes wider.

On the other hand, if the vehicle steering characteristics due to the suspension device change toward oversteer during rolling,
A wheel load transfer control device is attached to the front wheel side. Therefore, if we represent the total roll stiffness on the horizontal axis (with the up direction pointing to the right on the horizontal axis) and the vertical axis of the roll stiffness distribution (with the increase on the rear side pointing down on the vertical axis), we can express it in two-dimensional coordinates. It is possible to draw an equal steering characteristic line that slopes downward to the right.

When the wheel load transfer control device is operated, the steering characteristic is upward-sloping to the right through the characteristic point of the suspension device, compared to the downward-sloping characteristic when the wheel load transfer control device is installed on the rear wheel side. By including more equal steering characteristic lines, the control range of steering characteristics becomes wider.

[Example] Hereinafter, an example of the present invention will be described with reference to parts 1 to 4 of the attached drawings.
The present embodiment, which will be explained based on the drawings, was implemented regarding a hydraulic stabilizer provided on the rear wheel side as a wheeled cargo type movement control device.

In FIG. 1, 2L and 2R represent left and right wheels on the rear side of the vehicle, 4 represents a wheel support member, and 6 represents a vehicle body, respectively. 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 suspension device including a shock absorber 10 and a coil spring 12 is provided between the suspension link 8 and the vehicle body. Note that the same suspension device is also provided for the left and right wheels on the front side. In this embodiment, the steering characteristic of the entire suspension system is roll steer when rolling.

The understeer direction changes as the camber angle changes.

As shown in FIG. 1, a hydraulic stabilizer 14 is provided on the rear side of the suspension device described above.

This hydraulic stabilizer 14 includes an actuator section 14 provided between the left and right suspension links 8 and the vehicle body 6.
A, and a control section 14B that controls roll rigidity during turning by this actuator section 14A.

The actuator section 14A is a hydraulic cylinder 20L installed corresponding to the rear left and right wheels 2L2R.

In addition to 20R, throttle valve 22A22B that generates damping force
, accumulators 24A, 24B for accumulating pressure, and electromagnetic switching valve 25 for communication, and each of these elements serves as a hydraulic flow path for first piping 26A, 26B and second piping 28A, 2.
The structure is interconnected by 8B. Here, the hydraulic cylinder 20L, 2OR1 first pipe 26A, 2
6B and the second pipes 28A and 28B constitute an actuator.

Each of the hydraulic cylinders 2OL and 2OR includes a cylinder tube 20a, a slidable piston 20b that separates the inside of the cylinder tube 20a into an upper cylinder chamber U and a lower cylinder chamber, and a shaft fixed to the piston 20b. The piston rod 20c extends in both directions and is configured as a double acting type. The hydraulic cylinders 20L and 2OR having this structure each have a lower end of the piston rod 20c attached to the suspension link 8,
The upper end of the cylinder tube 20a is placed in a free state, and the end of the cylinder tube 20a is swingably supported by the vehicle body 6, whereby the hydraulic cylinders 2OL,
2ORs are respectively installed upright on the left and right sprung upper springs in parallel with the suspension device.

Further, the upper cylinder chamber U of the left wheel hydraulic cylinder 20L is connected to the lower cylinder chamber of the right wheel hydraulic cylinder 2OR via the first pipe 26A, and the upper cylinder chamber U of the left wheel hydraulic cylinder 20L is connected to the lower cylinder chamber of the right wheel hydraulic cylinder 2OR.
The lower cylinder chamber of the right-wheel hydraulic cylinder 2OR is connected to the upper cylinder chamber U of the right-wheel hydraulic cylinder 2OR via the first pipe 26B, thereby being cross-connected to each other. The first pipes 26A and 26B have tM! at their intermediate positions. They are interconnected via a switching valve 25. ! Magnetic switching valve 25
The solenoid has a normally open (communicating) structure that is in the "closed" position and the "open" position depending on whether the control signal S applied to the solenoid is turned on or off.

Furthermore, a second piping 26A, 26B has a second
Pipes 28A and 28B are connected to each other. The second pipes 28A, 28B are individually connected to accumulators 24A, 24B, and throttle valves 22A, 22B are individually inserted in the middle of the pipes 28A, 28B.

On the other hand, the control section 14B includes a control cylinder 30 that energizes the internal pressure of the actuator section 14A, and third pipes 32A and 3 connected to this control cylinder 30.
2B, an electric motor 34 that drives the control cylinder 30, and a controller 36.2B for controlling load movement. Vehicle speed sensor 38. and a steering angle sensor 39.

Of these, the control cylinder 30 is a double cylinder, similar to the aforementioned hydraulic cylinders 2OL and 2OR.

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.
. Among these, cylinder chamber R1,
R2 connects to the second pipe 28A via the third pipes 32A and 32B.
, 28B, respectively. Also, piston rod 3
One end of 0C is placed in a free state, and a rack 30 is placed on the other end.
d is formed. A pinion 34a of the electric motor 34 is engaged with this rack 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 V according to the vehicle speed to the controller 36. 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 V and θ from the vehicle speed sensor 38 and steering angle sensor 39 are inputted, 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 microcomputer of the controller 36 will be explained. The process shown in the figure starts when the power is turned on.

To explain this, in step (3) in the figure, the microcomputer of the controller 36 reads the detection signals V 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 (2). . In this step (2), well-known calculations (for example, JP-A-62-29316
7) is performed to estimate the lateral acceleration α7.

Thereafter, the process moves to step (2), where 1αvl>α, with respect to the preset standard lateral acceleration α7° near zero. 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 (2), the control signal S is turned off, and the rear electromagnetic switching valve 25 is opened or kept open.

On the other hand, if rYES is determined in step (2), control of wheel load movement is necessary, and the process proceeds to step (2). In this step ■, the control signal S is turned on,
The electromagnetic switching valve 25 is closed or kept closed.

Next, the process moves to step (2), and with reference to the map corresponding to FIG. 3 stored in the memory in advance, the lateral acceleration 1αY is calculated.
Motor rotation angle command value 1θN uniquely determined according to 1
Calculate 1. The command value θ, 41 characteristic shown in FIG. 3 is designed to increase as the lateral acceleration αV 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 rYEs in this determination of the steering direction, 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], the input signal θR is used to determine whether or not the electric motor 34 has rotated clockwise by the command value θ. 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 θN.

On the other hand, when determining rNOJ in step (2), steps @ to [phase] and (2) are performed in the same manner as steps (2) to (2). As a result, the electric motor 34 rotates to the left by the command value θi.

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 αa'-IO with the start of the process shown in FIG. 2, the control signal -OFF is maintained and the electromagnetic switching valve 25 is in the 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 steering characteristics determined by the suspension device can be obtained without causing load movement 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, since the estimated lateral acceleration α7 is almost zero, the switching valve 25 remains in communication, and the internal pressure of the actuator section 14A is not actively controlled, but is generated by each shock absorber 10. The damping force dampens the bounce. On the other hand, if the rear wheels 2L and 2R bounce due to passing through the convex portion, the hydraulic cylinder 2
When the strokes of 0L and 2OR are about to contract, the upper cylinder chambers U of the cylinders 2OL and 2OR are both simultaneously contracted, and the lower cylinder chambers are both simultaneously expanded. However, the compressed hydraulic oil in the cylinder chamber U is transferred to the pipe 2.
6A, 26B and the switching valve 25, almost equal amounts flow into the lower cylinder chamber of the same cylinder and the lower cylinder L of the opposite cylinder.
Since the changes in volume are the same since both cylinders are of the rond type, the amount of hydraulic oil overflowing from the cylinder chamber U is stored in the lower cylinder chamber with almost no difference, and the second pipe 28A.

There is no change in the amount of oil in 28B. This is exactly the same in the case of rebound due to passage through a recess, except that the vertical relationship is reversed.

Therefore, since the hydraulic oil does not pass through the throttle valves 22A and 22B during both bounce and rebound, almost no damping force is generated by the hydraulic stabilizer 14, and deterioration of riding comfort when traveling on an uneven road accompanied by bounce is prevented.

In addition, in such a straight-ahead state, one of the wheels 2L, for example,
Suppose that a stroke variation occurs due to riding over a protrusion, etc. In this case as well, there is no active internal pressure control of the actuator section 14A, but since the electromagnetic switching valve 25 is in communication,
The hydraulic oil overflowing from the cylinder chamber U flows into its own lower cylinder chamber via the switching valve 25. In other words, the switching valve 2
Due to the bypass action of No. 5, no change in oil amount occurs in the pipes 28A, 28B, and no damping force is generated. Therefore, even in the case of such one-wheel bounce, riding comfort is hardly impaired.

Now assume that the vehicle has transitioned from the above-mentioned straight-ahead state to, for example, turning left on a good road. This causes a load shift from the inner wheel to the outer wheel, and when viewed from the rear of the vehicle, the vehicle body on the right wheel 2R side sinks and the vehicle body on the left wheel 2L side tends to lift up. During this turn, the controller 36 estimates and calculates the lateral acceleration α1 of the sign corresponding to the left turn from the steering angle θ and the vehicle speed ■, and calculates the lateral acceleration 1αv 1
Motor rotation angle command values 1θ, 4 that are uniquely determined corresponding to
1 is set by manbu 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 (in the present example, counterclockwise in FIG. 1) by an angle of 00 minutes. Forced by this rotation, the piston rod 30c
It moves toward the right end in the figure by an amount corresponding to the rotation angle θi. As a result, one cylinder chamber R2 of the control cylinder 30 is compressed and the pressure in the intake R2 increases, while the other cylinder chamber R1 is expanded and the pressure in the intake R1 decreases.

As a result, the working pressures in the upper cylinder chamber U of the right hydraulic cylinder 2OR and the lower cylinder chamber of the left hydraulic cylinder 2OL rise simultaneously, and the working pressures in the cylinder chambers U and L of the opposite side rise simultaneously.
The working pressure of will drop at the same time.

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 20L and 2OR, and an axial force ΔF due to this pressure difference ΔP acts on the cylinder rod 20C. This axial force ΔF acts downward (direction toward the road surface) on the rear right side and upward (direction toward the vehicle body) on the rear left side, and its magnitude 1ΔF1 is almost the same. Therefore, the reaction force acting on the vehicle body is in the opposite direction to the above-mentioned direction.

Therefore, on the rear side, the moment due to the axial force ΔF acts in a direction that resists rolling, and the roll rigidity on the rear side increases. This allows the front side.

The total roll rigidity on the rear side also increases, the amount of load transferred from the inner turning wheel to the outer turning wheel increases, and the roll angle is kept small.

At this time, the roll rigidity distribution on the rear side increases, so the amount of load transfer on the rear side increases relative to the front side.
Vehicle steering characteristics are controlled in the oversteer direction. On the other hand, when the roll angle is small (suppressed), roll steer and camber angle changes occur in the suspension device, and the steering characteristics tend to oversteer.

Here, the magnitude of the control effect on the total steering characteristics will be examined based on the difference in the installation position of the hydraulic stabilizer 14.

In FIG. 4, the longitudinal axis shows the distribution of roll stiffness in the front and rear, and the horizontal axis shows the magnitude of the total roll stiffness in the front and back. As shown by the vertical axis in the figure, as the roll stiffness distribution increases toward the rear side, the steering characteristic changes toward oversteer (O3), and as it increases toward the front side, the steering characteristic changes toward understeer (US). Further, as shown by the horizontal axis in the figure, for the same lateral acceleration, as the total roll rigidity increases, the roll angle decreases, and as the total roll rigidity decreases, the roll angle increases. The steering characteristic of the suspension device of this embodiment is such that "when it rolls, it changes in the understeer direction." Therefore, as the roll angle increases and moves to the left of the horizontal axis, the steering characteristic changes in the understeer direction, and the roll angle increases. As it becomes smaller and moves to the right on the horizontal axis, the direction of oversteer changes.

Therefore, in the coordinates of the same figure, the total steering characteristic shifts toward oversteer as it advances toward the upper right, and toward understeer as it advances toward the lower left. As a result, the equal steering characteristic line (the steering characteristic The line formed by connecting equal points) is the dotted line C2, which is downward to the right.
..., denoted by C.

Let's superimpose the steering characteristic line A or B on this graph when the wheel load transfer control device (hydraulic stabilizer 14) is installed on the front side or the rear side and control is performed. When the wheel load transfer control device is installed on the front side, the increase in the front roll stiffness from the characteristic point O of the suspension device (that is, the increase in the distribution of roll stiffness on the front side and the increase in the total roll stiffness) causes A1 to decrease to the right. A line A2 is shown, which slopes upward to the left due to a decrease in front roll stiffness (that is, a decrease in the distribution of roll stiffness to the front side and an elimination of total roll stiffness). Line A is formed.

On the other hand, if the wheel load transfer control device is installed on the rear side,
An increase in the rear roll stiffness from the characteristic point 0 of the suspension device (that is, an increase in the distribution of roll stiffness to the rear side and an increase in the total roll stiffness) causes a line B, which slopes upward to the right, and a decrease in the rear roll stiffness ( In other words, a line Bt that slopes downward to the left is shown due to a decrease in the distribution of roll rigidity to the rear side and a decrease in total roll rigidity, so a line B that slopes upward to the right with respect to the horizontal axis is formed as a whole.

Therefore, when comparing the A characteristic line and the B characteristic line in the graph, the equal steering characteristic line C1.
..., the number of C lines is greater in the B characteristic line depending on its slope. Therefore, in the case of the vehicle characteristics of this embodiment, it can be seen that the control width of the steering characteristics is larger when the wheel load transfer control device is installed on the rear side.

Based on the above principle, in this embodiment, the hydraulic stabilizer 14 is installed on the rear wheel side, so it is possible to reliably prevent the situation where the steering characteristic control effect cannot be fully exerted, which would occur if the hydraulic stabilizer 14 was installed on the front wheel side. , a significantly larger total control width is obtained, and the maximum steering characteristic control effect is always exerted.

The above-mentioned control is the same when turning right. Additionally, if the pressure in the actuator section 14A changes suddenly during a turn due to unevenness on the road surface, this pressure fluctuation causes the throttle valve 22A to change suddenly.
, 22B.

In addition, contrary to the above explanation, if the steering characteristic of the suspension device installed alongside the wheel load transfer control device changes in the oversteer direction during rolling, the wheel load transfer control device should be installed on the front wheel side. This provides a significant advantage compared to installing it on the rear wheel side. In other words, in this case, the characteristic diagram corresponding to FIG. 4 becomes as shown in FIG. 5, and the one installed on the front wheel side (characteristic line A
) includes more of the equal steering characteristic line C than (characteristic line B) installed on the rear wheel side, and the control width of the steering characteristic becomes larger.

Further, the wheel load movement control device of the present invention may be implemented by, for example, a stabilizer device that utilizes the torsional rigidity of a torsion bar, in addition to the hydraulic stabilizer related to the X pipe as described above. Furthermore, active suspensions (for example, Japanese Patent Application Laid-Open No. 62-1992-1) change roll rigidity by separately providing fluid pressure cylinders between the wheels and the vehicle body and controlling the operating pressure of the fluid pressure cylinders based on turning information such as lateral acceleration. 295714) may be used as a wheel load movement control device.

Further, the structure of the actuator according to the present invention is not limited to that described in the above-mentioned embodiments. For example, the insertion directions of the hydraulic cylinders may be opposite to each other on the left and right sides, and for example, in FIG. The cylinder tube 20a of the hydraulic cylinder 2OR is attached to the suspension link 8, and the upper end side of the piston rod 20c is attached to the vehicle body 6), and the cylinder chambers U and L of both cylinders are apparently
It may also be a configuration in which they are connected in parallel;
It is possible to obtain the same effects as those of the embodiments described above.

Furthermore, the actuator in the present invention is for communication in the above-described embodiment! A structure may be adopted in which the value switching valve 25 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 explained above, in the present invention, when the wheel load transfer control device is installed on only one of the front and rear wheels, the steering characteristics of the suspension device to which the wheel load transfer control device is installed will be in the oversteer direction during rolling. If the steering characteristics of the suspension device changes in the understeer direction during rolling, the wheel load transfer control device is installed on the rear wheel side. The installation position takes into account the direction of change in steering characteristics due to the effects of roll steer and camber angle changes, so you can always obtain the maximum control width of steering characteristics under a given environment, and control steering characteristics. You can get the full effect.

[Brief explanation of drawings]

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 illustrating the control effect of the steering characteristic based on the total roll rigidity and roll rigidity distribution. FIG. 5 is an explanatory diagram of control effects corresponding to FIG. 4 according to another embodiment. The main symbols in the diagram are 6...Vehicle body, 8...Suspension link, 10...Shock absorber, 12...
・Coil spring, 14...hydraulic stabilizer,
14A... actuator section, 14B... control section,
20L, 2OR...hydraulic cylinder, 26A, 26B.
...first piping, 28A, 28B...second piping.

Claims (1)

[Claims] (1) Wheel load transfer control in which an actuator whose roll stiffness can be changed is inserted between either the left or right wheels at the front or rear of the vehicle and the vehicle body, and the roll stiffness of this actuator is changed to control the amount of movement of the wheel load. In the device, if the vehicle steering characteristic due to the suspension device provided between the four wheels and the vehicle body changes in the oversteer direction during rolling, one of the left and right wheels is set as the front wheel, and the vehicle steering characteristic due to the suspension device changes in the rolling direction. 1. A wheel load transfer control device characterized in that when a change occurs in an understeer direction, one of the left and right wheels is used as a rear wheel.