JPH04135910A - Wheel load shift control device - Google Patents

Abstract

PURPOSE:To enhance the maneuverability and stability upon turning with the use of a hydraulic stability or the like, by detecting or estimating a lateral acceleration in the vehicle width-wise direction, and causing an actuator to produce a roll rigidity which cancels a vibration in steering characteristic in accordance with the lateral acceleration. CONSTITUTION:A hydraulic stabilizer 14 composed of an actuator section 14A and a roll rigidity control section 14B is arranged on the rear side of a suspension device. Meanwhile, a controller 36 receives a vehicle speed V from a vehicle speed sensor 38, a steering angle theta from a steering angle sensor 39, and estimates a lateral acceleration alphar. Further, a solenoid selector valve 25 is closed if the estimated acceleration is greater than a reference value, and a motor rotating angle instruction value for canceling a variation in steering characteristic is calculated in accordance with the estimated acceleration alphar. Accordingly, a control cylinder 30 is driven by means of a motor 34 so that the internal pressure of the actuator section 14A is controlled.

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial Application Field] The present invention relates to an improvement of a wheel load movement control device such as a variable stabilizer that is inserted between a wheel and a vehicle body and is capable of changing roll stiffness.

[Conventional technology]

Conventionally, as a device for controlling the amount of movement of the wheel load, for example, the device described in Utility Model Application No. 60-76506 (the name of the device is "
"Hydraulic stabilizer") is known.

This conventional device has single-rod, double-acting hydraulic cylinders installed vertically between the left and right suspension arms of the vehicle 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.

Therefore, when the car body rolls, a pressure difference in the opposite direction occurs between the upper and lower cylinder chambers of the left and right hydraulic cylinders, which generates a force that resists the rolling and also causes the orifice to Damping force is generated based on the aperture effect. Furthermore, even when the vehicle body bounces, hydraulic oil passes through the orifice, generating a damping force.

[Problem to be solved by the invention]

However, the above-mentioned conventional device is configured to passively increase the roll rigidity in response to an increase in the lateral acceleration acting on the vehicle, and does not take into account the longitudinal distribution of the roll rigidity in response to the difference in lateral acceleration during turning. Although it is possible to suppress the roll angle during a turn, for example, in a vehicle designed to exhibit weak understeer characteristics, the larger the lateral acceleration, the more the traveling trajectory expands to the outside of the turn, and the tendency to understeer becomes stronger. There has been a problem in that the steering characteristics of the base vehicle (vehicle before the wheel load transfer control device is installed) fluctuate in response to changes in acceleration, resulting in a decrease in maneuverability and stability during turns.

The present invention was made in view of the problems of the conventional device, and the problem to be solved is to be able to maintain the steering characteristics of the base vehicle even when the lateral acceleration during turning is different; The goal is to maintain constant maneuverability and stability even when turning conditions vary.

[Means to solve the problem]

In order to solve the above problem, the claimed invention, as shown in FIG. A wheel load movement control device that controls the amount of movement of the load includes a lateral acceleration detection/estimation means for detecting or estimating lateral acceleration occurring in the vehicle width direction, and a lateral acceleration detection/estimation means according to a detected value or estimated value of the lateral acceleration detection/estimation means. Accordingly, roll stiffness control means is provided for causing the actuator to generate roll stiffness that cancels out changes in vehicle steering characteristics caused by turning conditions.

[Effect]

The roll stiffness control means operates in accordance with the detected lateral acceleration or the lateral acceleration estimated from physical quantities such as the steering angle and vehicle speed, and maintains the steering characteristics of the base vehicle as is.

For example, if the steering characteristic of the base vehicle before installing the wheel load transfer control device is understeer, an actuator is installed on the rear side, and the roll rigidity on the rear side is increased as the detected or estimated lateral acceleration increases. This increases the distribution of roll rigidity on the rear side relative to the front side, and the load shift from side to side on the rear side also increases.
Accordingly, the steering characteristics are changed in the direction of oversteer.

This changed steering characteristic substantially cancels out the change in the steering characteristic in the understeer direction due to lateral acceleration, so the initial understeer characteristic can be maintained as is. Further, when the steering characteristic of the base vehicle is oversteer, for example, the roll rigidity of an actuator installed on the front side is increased to increase the wheel load from the front inner wheel to the outer wheel. As a result, the change in steering characteristics caused by the generation of lateral acceleration and the change in the steering characteristics in the understeer direction due to the increase in the amount of front side load transfer are almost canceled out, and almost constant steering characteristics can be obtained. .

〔Example〕

Embodiments of the present invention will be described below with reference to the accompanying drawings.

(First Example) A first example is shown in FIGS. 2 to 6. In this embodiment, a hydraulic stabilizer as a wheel load movement control device is provided on the rear wheel side.

In FIG. 2, 2L and 2R represent the left and right wheels on the rear side of the vehicle, 4 represents a wheel support member, and 6 represents the 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 characteristics of the entire vehicle are designed to exhibit a predetermined understeer.

As shown in FIG. 2, a hydraulic stabilizer 14 is provided on the rear side of the above-mentioned suspension device.

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 for the rear left and right wheels 2L.

Hydraulic cylinder 20L equipped for 2R.

In addition to 20R, there is a throttle valve 22A that generates damping force.

22B1 Accumulators 24A and 24B for accumulating pressure, and T! for communication. It has a 1T11 switching valve 25, and each of these elements connects the first piping 26A, 26B and the second piping as a hydraulic flow path.
They have a structure in which they are interconnected by pipes 28A and 28B. Here, hydraulic cylinders 2OL, 2OR1 first
The piping 26A, 26B and the second piping 28A, 28B constitute an actuator.

Each of the hydraulic cylinders 20L and 20R includes a cylinder tube 20a, a slidable piston 20b that separates the interior 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 double-acting. 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 is placed in a free state, and the end of the cylinder tube 20a on the free end side is swingably supported by the vehicle body 6, whereby the hydraulic cylinders 2OL,
2OR is on the left and right.

Each of them is erected in parallel with the suspension device at the bottom of the spring.

Further, the upper cylinder chamber U of the left wheel side hydraulic cylinder 2OL is connected to the lower cylinder chamber of the right wheel side hydraulic cylinder 2OR via the first pipe 26A.
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. First piping 26
A and 26B are connected to each other via an electromagnetic switching valve 25 at an intermediate position. The electromagnetic switching valve 25 is turned on by the control signal S given to its solenoid.

Always open (closed or open when turned off)
It has a continuous structure.

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 a throttle valve 22A is provided in the middle of the pipes 28A, 28B.

22B are individually inserted.

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 has double cylinders, similar to the aforementioned hydraulic cylinders 2OL and 20R.

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 3Qb and extending in both axial directions. The cylinder chambers R1 and R2 are the third piping 32A.

It communicates with the second pipes 28A and 28B via 32B. 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 an 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 (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.

Here, the control principle of the present invention is shown in FIGS. 4(a) and (b).
Explain using. In the same figure (a), ω), the horizontal axis shows the front side left-right difference (relative difference) in the wheel load, and as the left-right difference moves from the zero origin 0 to the right side, the left wheel side wheel load increases, and the left side wheel load increases. As the vehicle progresses further, the wheel load on the right wheel increases. In addition, the vertical axis shows the difference (relative difference) in the rear side wheel load, and as the left and right difference moves upward from the zero origin 0, the wheel load on the left wheel increases, and as it moves downward, the wheel load on the right wheel increases. Wheel load increases.

Now, let us consider a case where the base vehicle has understeer characteristics and the wheel load transfer control device is provided only on the rear side.

In order to maintain the steering characteristics of the base vehicle when traveling straight (i.e., lateral acceleration α, -0), it is sufficient that the wheel load transfer control device is inactive and the characteristic point of the device is at the origin O. For example, suppose that the vehicle turns to the right from this state. During this right turn, as the generated lateral acceleration α7 increases, the understeer characteristic becomes stronger and the traveling trajectory expands toward the outer wheels.

Therefore, by using a wheel load transfer control device, roll stiffness is generated so that the wheel load on the rear left wheel (outer wheel) side is relatively larger than that on the rear right wheel (inner wheel) side, and the steering characteristics are shifted to the oversteer direction. let

That is, the characteristic point is set to y in Fig. 4(a) in proportion to the lateral acceleration.
+ (for example, when αY=0.IG), yz (for example, when αY=0.2G), y2 (for example, α,=0°3G)
), it is possible to correct the increase in understeer while increasing the overall roll rigidity. In addition, when turning left, the characteristic point is set to -3'1, for example, in FIG. '/z,
) '3.

In addition, if the base vehicle has understeering characteristics and the wheel load transfer control device is installed on the front side, the rear roll stiffness can be increased by weakening the front roll stiffness in proportion to the lateral acceleration α7. Although the roll stiffness can be relatively strengthened and the total roll stiffness decreases, the object of the present invention can be achieved. That is, when turning to the right, the characteristic point is set to, for example, −Xl+ in FIG.
xz, -X3, and when turning left, for example, XI l xi, X3.

On the other hand, it is assumed that the steering characteristic of the base vehicle is oversteer and the wheel load movement control device is provided on the front side. In this case, the degree of oversteer changes depending on the change in lateral acceleration α, so in order to correct this and maintain the initial steering characteristics, the roll stiffness during turning should be adjusted in proportion to the magnitude of lateral acceleration α7. , so that the difference in load between the inner and outer rings becomes larger on the outer ring side. That is, when turning to the right, the characteristic point of the wheel load transfer control device is set to, for example, XI t Xt + X3 in FIG.
f, which also increases the overall roll stiffness.

Further, it is assumed that the steering characteristic of the base vehicle is oversteer and the wheel load transfer control device is provided on the rear side. In this case, the roll stiffness on the rear side should be weakened in proportion to the magnitude of the lateral acceleration αV, and the front and rear distribution of roll stiffness should be relatively increased on the front side.In other words, the characteristic point of the wheel load transfer control device should be set to the right. When turning, for example -)'+ in Figure (a)
, )'z.

y3, and when turning left, for example, V+, 3'z, Y3. Although this reduces the overall roll stiffness,
Understeer is achieved in response to changes in lateral acceleration α7.

Next, the process shown in FIG. 3 executed by the microcomputer of the controller 36 will be explained based on the above-mentioned principle. 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
(See the method shown in Publication No. 7) to estimate the lateral acceleration α.

Thereafter, the process moves to step (3), and it is determined whether 1αyl>α7° with respect to a preset standard lateral acceleration α7° near zero. If the result of this judgment is "NO", 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, when it is determined in step (2) that it is rYES, it is determined that 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 (3), and with reference to the map corresponding to FIG. 5 stored in the memory in advance, the lateral acceleration 1α71
Motor rotation angle command (ltlθ8
Calculate 1. The command value θ, 1 characteristic in Fig. 5 suppresses the roll angle and keeps the steering characteristic constant.
Lateral acceleration 1α.

It is expected to increase with the increase in

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 rYEsJ in this determination of the steering direction, steps ① to 0 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. 2) in step (2). 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 rYESJ, the motor rotation is stopped at 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 rNOJ in step (2), the processing in steps @ to 0.0 is performed in the same manner as in steps (2) to (2). As a result, the electric motor 34 rotates to the left by the command value θ9.

Among the above explanations, vehicle speed sensor 38. Steering angle sensor 39
．． The processing of steps ① and ② in Fig. 3 constitutes a lateral acceleration detection/estimation means. Also, the process of steps ① to 0 in FIG. 3, electric motor 34. Control cylinder 30. and third piping 32A.

32B constitutes roll rigidity control means.

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

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

Since the controller 36 estimates the lateral acceleration α7ζ0 upon starting the process shown in FIG. 3, the control signal is maintained at OFF and the electromagnetic switching valve 25 is 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, the initial understeer characteristics of the base vehicle are maintained without wheel load movement occurring 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 lO. 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, 2. When the strokes of the hydraulic cylinders 20L and 2OR are about to contract, the upper cylinder chambers U of the cylinders 2OL and 20R are simultaneously reduced, and the lower cylinder chambers are simultaneously expanded. However, the compressed hydraulic oil in the cylinder chamber U flows through the pipes 26A, 26B and the switching valve 25 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 the same since both cylinders are of the Rondo type, the 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, 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 α7 whose sign corresponds to the left turn from the steering angle θ and the vehicle speed V, and maps the motor rotation angle command value 1θH which is uniquely determined corresponding to the lateral acceleration 1α71. Set by 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 (counterclockwise in FIG. 2 in the present example) by an angle of 09 minutes. Forced by this rotation, the piston rod 30c moves to the second position.
It moves toward the right end in the figure by an amount corresponding to the rotation angle θ. 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 of 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 cylinder chamber U of the opposite side increases.
The operating pressure of L drops to the knees 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 2OL 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. As a result, the roll stiffness distribution on the rear side increases, the amount of load transferred from the inner wheel to the outer wheel on the rear side increases, the steering characteristics of the vehicle are changed to a neutral steer direction, and the driving trajectory expands to the outside of the turn. This not only makes it possible to reliably correct the problem, but also increases the total roll rigidity of the vehicle and suppresses the roll angle. And the lateral acceleration 1α
As 71 becomes larger, the amount of internal pressure control for the actuator section also becomes larger, so the amount of correction of the steering characteristics is also adjusted in response to the change in the lateral acceleration α7, and the initial steering characteristics of the base vehicle are maintained almost as they are. This eliminates the situation where maneuverability and stability change even if the turning speed changes.

The roll stiffness control described above is the same when turning right. Furthermore, 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 spring force of the accumulators 24A and 24B and the throttle valve 22A to change rapidly.
, 22B.

FIG. 6 qualitatively shows the state of the control at a certain point in time described above. In the figure, vector v0 indicates the amount of load movement on the front side and rear side of the base vehicle including the suspension device, and vector vr indicates the amount of load movement on the side of the hydraulic stabilizer 14.

In addition, contrary to the above explanation, if the initial steering characteristic of the base vehicle equipped with a wheel load transfer control device such as a hydraulic stabilizer is oversteer, the wheel load transfer control device should be used as in the control principle described above. Installed on the front wheel side, lateral acceleration 1α.

The front roll stiffness may be controlled to increase as the front roll stiffness increases. As a result, the share of roll stiffness on the front side increases in proportion to the magnitude of the lateral acceleration, and the steering characteristics are changed in the direction of understeer. Therefore, even if the initial oversteer characteristic changes due to a change in lateral acceleration, the amount of change is canceled out and the predetermined oversteer characteristic is maintained. Further, the total roll stiffness of the vehicle also increases in accordance with the lateral acceleration, and the roll angle is maintained at a small value. Furthermore, since the installation position of the wheel load transfer control device can be changed depending on the difference in the steering characteristics of the vehicle, there is an advantage that the range of selection is widened and the degree of freedom in vehicle design is improved.

(Second Embodiment) Next, a second embodiment will be described with reference to FIGS. 7 to 10. Here, the same components as in the first embodiment are denoted by the same reference numerals, and the description thereof will be omitted or simplified.

In the second embodiment, the actuator section 14A of the hydraulic stabilizer 14 as a wheel load movement control device is placed on the front side,
These actuator parts 14A are installed individually on the rear side.
, 14A can be controlled independently.

To explain this using Figure 7, 2FL and 2PR are the front left wheel and front right wheel, and 21? L and 2RR are the rear left wheel and rear right wheel. Front wheel 2FL, ZFR side and rear wheel 2RL
, 2RI? The same actuator parts 14A as in the first embodiment are installed on the sides, and control cylinders 30 are separately provided corresponding to these actuator parts 14A, 14A. It is driven in the same way as in the first embodiment. Each of the electric motors 34 is supplied with a motor drive signal i from a common controller 36 and is driven separately. The other configurations are the same as the first embodiment.

Here, the control principle of the second embodiment will be explained with reference to FIG. The vertical and horizontal axes in this figure are plotted in the same manner as in FIG. 4. Now, assuming that the steering characteristic of the base vehicle is a predetermined understeer, the control line that corrects the change in the steering characteristic due to the change in lateral acceleration α7 when the wheel load transfer control device is installed only on the rear side is as follows. As mentioned above, it moves on the vertical axis, and when the control device is installed only on the front side, it moves on the horizontal axis as described above. (The direction in which the left and right load difference exists is the same control direction.) Therefore, a control line that maintains the steering characteristic constant when the lateral acceleration α changes can be drawn as shown by the dotted line in the figure. On the other hand, the turning situation
In other words, in order to keep the roll angle constant when the lateral acceleration α7 changes, it is sufficient to keep the total roll rigidity constant regardless of whether the wheel load transfer control device is installed on the rear side or the front side. , the control line shown by the dashed line in the figure can be drawn in a state where it intersects with the control line with constant steering characteristics.

Therefore, when trying to keep both the steering characteristic and the roll angle constant when the lateral acceleration α7 changes, a line is formed by connecting the points where both control lines intersect for each same value of the lateral acceleration α. , the control may be performed along the control line A shown in solid line.

A specific example based on this control line A will be given. Now, the lateral acceleration αY=0. When using IG, when the wheel load transfer control device is installed only on the rear side, the required control amount to maintain a constant roll angle is y
, and the necessary control amount to maintain constant steering characteristics is yb, and it is not possible to satisfy both at the same time. On the other hand, if the control device is provided on both the rear side and the front side, the control amount yc on the rear side and the control amount y on the front side may be used. In other words, the required amount for keeping the roll angle constant is "yc+yaJ" since the control directions are the same before and after.
(=ym), and since the control direction is opposite in the front and rear, the necessary amount to maintain a constant steering angle is ``)'CVa J (
=3'b), and both the constant steering characteristic and constant roll angle are satisfied at the same time.

Note that when the base vehicle exhibits oversteer characteristics, the same control is performed based on the control line formed by interchanging the vertical and horizontal axes in FIG. 8.

Based on the above principle, the controller 36 performs control according to the same flowchart as in FIG. 3 showing the first embodiment. However, in the process corresponding to step ■ in Figure 3, step 9
Referring to a map in which the characteristics shown in the figure are stored in advance, the rear side command value θH7 and the front side command value θ□ are individually set. The characteristics shown in FIG. 9 reflect the control amount determined by the control line A in FIG. 8 described above.

Therefore, when turning, the internal pressures of both the rear and front actuator sections 14A, 14A are controlled individually, and the direction of internal pressure control is opposite to maintaining constant steering characteristics, and the roll angle is controlled in the opposite direction. It is consistent for holding constant. As a result, it is possible to obtain a wheel load movement amount (for example, yb in Fig. 8) that offsets the change in steering characteristics caused by the lateral acceleration α7 at that time, and also to maintain a constant roll angle. The amount of load movement (for example, y in FIG. 8) is obtained. This load movement control is performed by constantly monitoring the value of lateral acceleration α7. Therefore, even if the value of the lateral acceleration α7 changes due to changes in turning conditions, the steering characteristics are kept almost the same and the roll angle is also kept constant.

FIG. 10 qualitatively shows the state of control at a certain point in time. In the figure, vector v0 indicates the amount of load movement on the front side and rear side of the base vehicle including the suspension device, and vectors vf and vr indicate the amount of load movement on the front side and rear side by the hydraulic stabilizer 14, respectively. .

Other operations are similar to those in the first embodiment.

Note that 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 toshi shim 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.

Furthermore, the structure of the actuator in the present invention is not limited to that described in the above-described embodiments. For example, the insertion direction of the hydraulic cylinders may be opposite to each other on the left and right sides, and for example, in FIG. 2, the hydraulic pressure on the rear right wheel side is (The cylinder tube 20a of the cylinder 20R 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 may be apparently connected in parallel. Also, the same effects as those of the above-mentioned embodiment can be obtained.

Furthermore, the actuator according to the present invention may adopt a structure in which the electromagnetic switching valve 25 for communication in the above-described embodiment is omitted.

Furthermore, in addition to the above-described configuration, the lateral acceleration detection/estimation means of the present invention may have a configuration that directly uses a lateral acceleration sensor that outputs a signal according to the inertial force and the direction of its action.

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, the present invention includes a turning state detection means for detecting a lateral acceleration occurring in the vehicle width direction or a physical quantity corresponding to the lateral acceleration, and a turning state detection means for detecting a lateral acceleration generated in the vehicle width direction or a physical quantity corresponding to the lateral acceleration, and Since the actuator is provided with roll stiffness control means for generating roll stiffness in a direction that cancels out changes in the vehicle steering characteristics, the steering characteristics of the vehicle before the wheel load transfer control device is actuated are, for example, understeer. Even if the understeer characteristics become stronger as the lateral acceleration increases, the change in the characteristics is almost offset by the control in the oversteer direction based on the roll stiffness changed by the actuator, and the steering characteristics always remain almost constant. Become.

Therefore, it is possible to reliably prevent the steering characteristics from changing due to differences in turning conditions, and eliminate fluctuations in maneuverability and stability due to differences in turning conditions.

[Brief explanation of the drawing]

FIG. 1 is a diagram corresponding to claims of the present invention. FIG. 2 to FIG. 6 are diagrams showing the first embodiment of the present invention;
The figure is a schematic configuration diagram, Figure 3 is a flowchart showing an example of processing in the controller, Figure 4 (a) ■) is an explanatory diagram showing the control principle, Figure 5 is a graph showing command value characteristics, and Figure 6 is a flowchart showing an example of processing in the controller. FIG. 2 is an explanatory diagram showing a state of control in a vectorial manner. 7 to 10 are diagrams showing a second embodiment of the present invention, in which FIG. 7 is a schematic configuration diagram, FIG. 8 is an explanatory diagram showing the control principle, and FIG. 9 is a diagram showing command value characteristics. The graph in FIG. 10 is an explanatory diagram showing the state of control in a vectorial manner. Main symbols in the diagram are 6...Vehicle body, 8...Suspension link, 14...Hydraulic stabilizer, 14A.
...Actuator section, 14B...Control section, 2OL2
OR... Hydraulic cylinder, 26A, 26B... First piping, 28A, 28B... Second piping, 30... Control cylinder, 32A, 32B... Third piping, 34
...Electric motor, 36...Controller, 38...
- Vehicle speed sensor, 39... Steering angle sensor. Figure 4: Lateral acceleration 1 Y 70/t Control wheel weight movement

Claims (1)

[Claims] (1) In a wheel load movement control device that includes an actuator whose roll rigidity can be changed between the vehicle wheels and the vehicle body, and controls the amount of movement of the wheel load by changing the roll rigidity of this actuator, in the vehicle width direction. lateral acceleration detection/estimation means for detecting or estimating the lateral acceleration generated in the lateral acceleration; and roll stiffness for canceling changes in vehicle steering characteristics caused by turning conditions according to the detected value or estimated value of the lateral acceleration detection/estimation means. and roll stiffness control means for causing the actuator to generate a wheel load movement control device.