JPH0492706A - Wheel load movement control device - Google Patents

Abstract

PURPOSE:To prevent worsening of steer characteristics and the occurrence of unstable cornering due to unexpected levitation of an inner wheel during cornering by correcting command value characteristics regarding movement of loads on both sides such that the more a vehicle weight or a loading weight is decreased, the more a cargo movement amount is decreased. CONSTITUTION:An actuator A roll rigidity of which is variable is located between a car body and wheels and through variation of the working state of the actuator A, a movement amount of a wheel load is controlled. In this device, a cornering state detecting means B to detect a cornering state is provided, and through collation of the detecting signal with command value characteristics regarding cargo movement between wheels on both sides by means of a command value setting means C, a command value for cargo movement is set. Working of the actuator A is controlled by an actuator control means D according to the set command value. A weight detecting and estimating means E to detect or estimate the weight or the loading weight of a vehicle is provided, and command value characteristics are corrected by a command value characteristic correcting means F so that the more a detecting value or an estimating value is decreased, the more a load movement amount is decreased.

Description

DETAILED DESCRIPTION OF THE INVENTION (Industrial Application Field) The present invention relates to a wheel load transfer control device, and particularly to a control device that appropriately controls load transfer between wheels caused by rolling.

[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-ended, double-acting, 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 chambers of each hydraulic cylinder are communicated in a cross state through hydraulic piping, 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 rolls, the cross-circulating hydraulic fluid generates a force that resists the roll, and the orifice's throttling effect generates a damping force.

(Problems to be Solved by the Invention) However, the conventional device described above has the following disadvantages because it has a configuration in which the amount of load transfer between the left and right wheels is simply controlled according to the lateral acceleration acting on the vehicle. In other words, when the steering characteristics are set to be optimized for a certain vehicle weight, for example, if the rear wheel roll stiffness is increased to increase the amount of load transfer in order to make the steering characteristics oversteer, the rear inner wheel In a situation where the rear inner wheel does not lift up, the steering characteristics can be controlled as desired. However, if the amount of load transfer increases and the amount of load transfer becomes larger than the initial load of the rear inner ring, the rear inner wheel will lift up, and the steering characteristics will change. will get worse.

In particular, if the rear wheels are the driving wheels, driving force will be lost.

The present invention was made in view of the problems of the conventional devices, and the problem to be solved is that even when controlling the amount of movement of the wheel load during turning, it is difficult to avoid problems caused by excessive load movement. This is to prevent the inner ring from lifting up and avoid deterioration of steering characteristics when turning.

[Means to solve the problem]

In order to solve the above problems, the invention according to claims (1) and (2), as shown in FIG. A wheel load movement control device that controls the amount of movement of a wheel load by changing the operating state includes a turning state detection means for detecting a signal indicating a turning state, and a detection signal of the turning state detection means for detecting a load between the left and right wheels. A command value setting means for setting a command value for load movement by checking with command value characteristics regarding movement, and an actuator control means for controlling the operation of the actuator according to the command value set by the command value setting means. , weight detection/estimation means for detecting or estimating the weight or loaded weight of the vehicle, and correcting the command value characteristic so that the smaller the detected value or estimated value of the weight detection/estimation means, the smaller the load movement amount. A command value characteristic correction means is provided. Among these, the command value characteristic correction means includes a configuration that advances the saturation timing of the command value characteristic as the detected value or estimated value of the weight detection/estimation means becomes smaller.

Further, as shown in FIG. 1(b), the invention according to claim (3) includes an actuator that can change the roll stiffness between the vehicle body and the wheels, and changes the operating state of the actuator to reduce the wheel load. In a wheel load movement control device that controls the amount of movement, a turning state detection means detects a signal indicating a turning state, and the detection signal of this turning state detection means is compared with a command value characteristic regarding load movement between the left and right wheels. A command value setting means for setting a command value for load movement, and an actuator control means for controlling the operation of the actuator according to the command value set by the command value setting means, and detecting the weight of the vehicle or the loaded weight. Alternatively, a weight detecting/estimating means for estimating the weight, and a control prohibiting means for prohibiting the operation of the actuator controlling means when the detected value or estimated value of the weight detecting/estimating means becomes a predetermined value or less are provided.

[Effect]

In the invention described in claim (1), the command value characteristic correction means comprises:
The command value characteristics are corrected so that the smaller the vehicle weight or loaded weight detected or estimated by the weight detection/estimation means, the smaller the command value is. On the other hand, since a signal corresponding to the turning state such as the steering angle and vehicle speed is detected by the turning state detection means, the command value setting means refers to the command value characteristic corrected by the command value characteristic correction means and detects the turning state. Set a command value corresponding to the detected value of the means. In other words, even in a turning state with the same lateral acceleration, the smaller the vehicle weight or loaded weight, the more
A small command value is set, and the actuator control means
The operation of the actuator is controlled based on this small command value.

Since the initial load of the wheels is smaller as the vehicle weight or loaded weight is smaller, in situations where the same lateral acceleration is occurring, the amount of load movement tends to exceed the initial load, and the inner wheel of the turn tends to lift up. However, this claim (1
In the invention described in ), when the vehicle weight or the loaded weight is small, the roll rigidity is made smaller than when the vehicle weight or the loaded weight is large, and the amount of load movement is suppressed, so that lifting of the inner ring is prevented.

In the invention set forth in claim (2), the command value characteristic correction means performs a correction that accelerates the saturation of the command value characteristic as the vehicle weight or loaded weight is smaller, so that the saturation start point is a signal indicating a turning state. Move to the smaller side. This results in
As the vehicle weight or loaded weight decreases, the command value tends to reach the maximum value, and roll rigidity saturates earlier when lateral acceleration or the like increases. Therefore, the amount of load movement between the left and right wheels is also suppressed compared to when the weight is large, and lifting of the inner ring is prevented.

In the invention described in claim (3), when the detected value or estimated value of the weight detection/estimation means becomes a predetermined value or less, the control prohibition means prohibits the operation of the actuator control means. Under conditions where the vehicle weight or loaded weight is reduced and the inner ring is likely to lift, the active operation of the actuator is discontinued, and therefore active load transfer control between the left and right wheels is discontinued, causing the inner ring to lift. Prevented.

〔Example] Embodiments of the present invention will be described below with reference to the drawings.

(First Embodiment) A first embodiment, which is an embodiment of the invention as claimed in claim (1), will be described based on FIGS. 2 to 4 of the accompanying drawings. Book 1
The example was carried out regarding a hydraulic stabilizer as a wheel load movement control device.

In Figure 2, 2L and 2R represent the left and right wheels of the vehicle, respectively.
6 indicates a wheel support member, and 6 indicates a 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 shore absorber 1 is provided between this suspension link 8 and the vehicle body.
0 and a coil spring 12 are provided.

Furthermore, in FIG. 2, 14 indicates a hydraulic stabilizer. This hydraulic stabilizer 14 includes a stabilizer main body 14A provided between the left and right wheel suspension links 8 and the vehicle body 6, and a control section 14B that controls the roll rigidity of the stabilizer main body 14A when turning.

Note that the left and right wheels 2L and 2R represent rear wheels, and although not shown, the front wheels also have the same suspension and stabilizer structure.

The stabilizer main body 14A includes hydraulic cylinders 2OL and 2OR installed corresponding to the left and right wheels, as well as a throttle valve 22A,
22B and accumulator 24A.

24B, each of which serves as a conduit for first hydraulic piping 26A, 26B and second hydraulic piping 28A, 28.
The structure is interconnected by B. Here, the hydraulic cylinders 2OL, 2OR and the first hydraulic pipe 26
A and 26B constitute actuators.

Each of the hydraulic cylinders 20L 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 has a piston rod 20c extending in the lateral direction, 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, thereby allowing the hydraulic cylinders 20L,
20R is the left and right wing.

They are set up in each section.

Then, the upper cylinder chamber U of the left wheel side hydraulic cylinder 2OL
is connected to the right wheel side hydraulic cylinder 2 via the first hydraulic pipe 26A.
The lower cylinder chamber of the left wheel hydraulic cylinder 20L is connected to the upper cylinder chamber U of the right wheel hydraulic cylinder 20R via the first hydraulic piping 26B. They are cross-connected to each other. Further, second hydraulic pipes 28A, 28B are connected to intermediate positions of the first hydraulic pipes 26A, 26B, respectively. The second hydraulic pipes 28A, 28B are connected to accumulators 24A, 24B, respectively, and throttle valves 22A, 22B are individually interposed in the middle of the pipes 28A, 28B.

On the other hand, the control section 14B includes a control cylinder 3 for movement control that controls the internal pressure of the stabilizer main body 14A.
0 and a third cylinder connected to this control cylinder 30.
hydraulic piping 32A, 32B and control cylinder 3
0, a controller 36, and a vehicle speed sensor 38. Steering angle sensor 39. and wheel load sensors 40a to 40d.

Of these, the control cylinder 30 has two rods, similar to the aforementioned hydraulic cylinders 20L 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.
, R2 and a slidable piston 30b, and a piston rod 30c fixed to the piston 30b and extending in the axial direction. Among these, cylinder chamber R1,
R2 communicates with second hydraulic pipes 28A and 28B via third hydraulic pipes 32A and 32B, 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. This rack 30d
A pinion 34a of the electric motor 34 is engaged with the pinion 34a of the electric motor 34.

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 composed of a pulse detector mounted on the steering shaft, and outputs a pulse signal θ according to the steering direction and steering angle to the controller 36. On the other hand, the wheel load sensors 40a to 40d are connected to the hydraulic cylinders 20L and 20R of the front and rear wheels and the vehicle body 6, respectively.
and a voltage signal W corresponding to the load carried by each wheel.
1 to W4 are output to the controller 36.

In this embodiment, the controller 36 includes a microcomputer, a motor drive circuit, a solenoid drive circuit, etc.
Vehicle speed sensor 38. The detection signals ■, θ, and WI-W4 from the steering angle sensor 39 and the wheel load sensors 40a to 40d are input, the processing shown in FIG. It looks like this. Note that a rotation angle sensor (not shown) is attached to the electric motor 34, and a motor rotation position signal θ from this sensor is attached to the electric motor 34.
R is supplied to the controller 36 and used to control the rotational position of the motor.

Next, the operation of the first embodiment will be explained.

First, the process shown in FIG. 3 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 controls the m load sensor 40a.
-40d detection signals W1 to W4 are read and the values are stored as wheel loads. Next, the process moved to step (2), and the wheel loads W1 to W4 read in step (2) were added to calculate the weight W (-W, +Wz +W3 +Wa) of the entire vehicle at that point, including the live load and the weight of the occupants. rear,
Move to step ■.

In step (2), the map of drive motor rotation angle (θM) characteristics is corrected in accordance with the vehicle weight W calculated in step (2). This map shows the lateral acceleration 1α71 as shown in Figure 4.
The drive motor rotation angle IθM increases as 1
(two command values), and the curve for the rotation angle 1θH1 has a rate of change that slows down as the lateral acceleration lα91 increases. The characteristic curve is prepared in advance in a state in which Iθ, 41 is gradually shifted to the increasing side as the vehicle weight W increases, and the map correction is performed on the multiple rotation angles 1θ, 41 curve (in the figure). W
This is done by selecting a curve corresponding to the vehicle weight W from among the following curves: = WH time, W = W L time (W) l > WL ). Therefore, as shown in FIG. 4, as the total sum W of the wheel loads W1 to W4 increases, the value of the drive motor rotation angle ]θ,1 corresponding to the same lateral acceleration 1αY1 increases.

Next, the controller 36 moves the process to step (2), reads the detection signals (2) and θ from the vehicle speed sensor 38 and the steering angle sensor 39, stores the values as the vehicle speed and steering angle, and then moves to step (2). In step ■,
The lateral acceleration α7 is estimated from the read values ■ and θ in step ■ by a well-known calculation (see, for example, the method disclosed in Japanese Patent Laid-Open No. 62-293167).

Further, the process moves to step ■, and the map of the drive motor rotation angle 1θH1 characteristics corrected in step ■ is referred to, and the rotation angle IθM 1 that is uniquely determined corresponding to the lateral acceleration 1α estimated in step ■ is read. After that, move on to step ■.

In step (2), it is determined whether the steering wheel is turned to the right from the sign of the steering angle signal θ input in step (2). If rYES in this judgment, step ■ ~ [phase]
Process. That is, the microcomputer outputs the motor drive signal j in the direction corresponding to the right rotation of the motor (clockwise direction in FIG. 2) in step (3). Next, in step (2), the motor rotational position signal θR is input, and in step (2), using the input signal θR, it is determined whether the electric motor 34 has rotated in the right direction by a command value of 68 minutes. and,
If "NO", repeat the process of step (2) and [phase], and if rYEsJ, stop the motor rotation in step (2), and then return to step (2). As a result, the electric motor 34 rotates in the right direction by the command value θ9.

On the other hand, when the determination in step (2) is "NO", 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 θ8.

In this embodiment, the vehicle speed sensor 38. The steering angle sensor 39 and the processing in steps (2) and (2) in FIG. 3 constitute a turning state detection means, and the wheel load sensors 40a to 40d and the processing in steps (2) and (2) in FIG. 3 constitute a weight detection means. Further, the process of step ① in FIG. 3 corresponds to the command value setting means, and the third
The process of step (2) in the figure corresponds to the command value characteristic correction means. Further, the processing of steps 1 to 0 in FIG. 3, the third hydraulic pressure distribution f32A, 32B, the control cylinder 30, and the electric motor 34 constitute actuator control means.

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

Upon starting the process shown in FIG. 3, the controller 36 first performs the following steps.
The vehicle weight W is detected by the detection signal W of the wheel load sensors 40a to 40d.
Obtained from 1 to W4. Then, since the characteristic curve data shown in FIG. 4 corresponding to the size of the vehicle weight W is read into the map storage area of the memory, the larger the vehicle weight W, the more the characteristic that connects the larger value of the drive motor rotation angle θ, 41. Set.

Now, suppose that the vehicle is traveling straight on a smooth road with a large load and data corresponding to the characteristic curve on the heavy side (W=WH) in Figure 4 is stored in the map. .

In this straight-ahead state, the lateral acceleration α estimated and calculated by the controller 36 is zero, so the drive motor rotation angle 1θM 1
=0 is set (see Figure 4).

Then, the controller 36 performs the step @ shown in FIG.
Although the processing of ~[phase] is performed, since the rotation angle is 1θ and l=0, no rotation command is given to the electric motor 34. As a result, the piston position 30b of the control cylinder 30 maintains the neutral position, and the internal pressure of the hydraulic stabilizer 14 is not energized.

Furthermore, since the current road surface is good and has almost no irregularities, the wheels 2L and 2R do not bounce or rebound. For this reason, the left and right hydraulic cylinders 20L.

2OR stroke change did not occur and piping was 26A.

No flow of hydraulic oil occurs within 26B, 28A, and 28B.

Therefore, the throttle valves 22A, 22B and the pipes 26A, 2
No damping force is generated due to the flow path resistance of 6B, 28A, and 28B, and predetermined suspension characteristics are maintained.

Assume that the vehicle shifts from the straight-ahead state described above to a turning state on a good road. Assume that this turning is, for example, a right turning, and rolling occurs in a direction in which the left wheel 2L side sinks and the right wheel 2R side rises (see arrow A in FIG. 2) when viewed from the rear of the vehicle. During this turn, the controller 1-roller 36 estimates and calculates the lateral acceleration α of the sign corresponding to the right turning direction, and the motor rotation angle lθI, which is uniquely determined corresponding to the lateral acceleration 1αY], is calculated as the fourth It is set by referring to a map corresponding to the curve when W=W in the figure. Now, αv
]=α, then θi 1− as shown in Figure 4.
6M2 is set.

Then, the controller 36 determines right turn steering based on the sign of the steering angle signal θ, and rotates the electric motor 34 to the right (clockwise in FIG. 2 in the present example) by an angle θ. Forced by this rotation, the piston rod 30c rotates by an angle θ toward the left end in FIG. Move by the amount corresponding to .

As a result, one cylinder chamber R1 of the control cylinder 30 is compressed, and the pressure in the cylinder chamber R1 increases, and at the same time, the other cylinder chamber R2 is expanded, and the pressure in the cylinder chamber R2 decreases. As a result, in each of the left and right hydraulic cylinders 2OL and 2OR, the upper cylinder chamber U
Since a pressure difference of ΔP is generated between the lower cylinder chambers, a force ΔF corresponding to this pressure difference ΔP acts on the piston rod 20c.

In this example, the direction of this force ΔF is downward (toward the road surface) at the left wheel hydraulic cylinder 2OL and upward (towards the vehicle body) at the right hydraulic cylinder 20R, so the reaction force acting on the vehicle body is A in the figure of the car body is opposite to the direction of
A moment is created that opposes the rolling in the direction. This proactively and proactively suppresses rolling,
The roll angle is controlled to a small value. Along with this, the second
The pressure rises in the hydraulic pipes 28A, 28B and is energized by the lower valve, and the accumulator 24 is increased through the throttle valves 22A, 22B.
Hydraulic oil flows between A and 24B. For this reason,
A damping force is generated due to the throttling effect of the throttling valves 22A and 22B, thereby damping vibrations.

The controller 36 repeatedly calculates the lateral acceleration αY,
Since the above-mentioned control is repeated according to the calculated value, the opposing moment is finely adjusted according to rolling changes during turning. As a result, the roll stiffness during turning is accurately controlled in cooperation with the damping effect of the shock absorber 10.
When the turning is completed and the vehicle returns to the straight-ahead state, it automatically returns to the neutral state described above.

In the case of a left turn, the above-mentioned operations are the same, although the left and right directions are reversed.

On the other hand, it is assumed that the weight W of the vehicle becomes lighter due to work such as stopping at -degrees and unloading loaded cargo, so that, for example, W=WL (<WH).

In such a case, the vehicle weight W = WL at that time is immediately detected, and the characteristics of the drive motor rotation angle 1θ, j are corrected according to the change in the weight W (steps
■). In the current example, the data corresponding to the lower curve when W=WL in Figure 4 is loaded into the map, so
For example, in the case of estimated lateral acceleration 1α, 1-1αYll,
Drive motor rotation angle Iθ, l=lθNi1 (<lθ, 21
). In other words, when stacking, the absolute value of the rotation angle, which is the command value for moving the load, is automatically adjusted to a smaller value than when stacking.

Therefore, even when transitioning to the turning state as in the case of stacking described above, the angle 1θM1 at which the controller 36 rotates the electric motor 34 becomes smaller, and the amount of movement of the piston 30b of the control cylinder 30 becomes smaller. Due to this reduction in the amount of movement, the differential pressure ΔP between the upper and lower cylinder chambers U and L in the left and right hydraulic cylinders 20L and 20R decreases,
That is, since the axial force ΔF and the damping force are reduced, the roll rigidity is automatically reduced compared to when stacked (W=WH), and as a result, the amount of load transfer between the left and right wheels is also automatically reduced.

In this way, in this embodiment, even in a driving situation where the same lateral acceleration α occurs, as the vehicle weight W decreases, the amount of load transfer between the left and right wheels decreases, so that the amount of load transfer between the inner wheels decreases. Since a situation in which the initial load is exceeded almost never occurs, lifting of the inner ring is accurately prevented.

Normally, the initial load on the wheels is small during heavy loading (W = WL), so if the roll rigidity is increased too much in favor of attitude control, the amount of load movement will easily exceed the initial load.
On the contrary, it may cause the inner race to lift up, resulting in deterioration of the steering characteristics. However, in this embodiment, as the vehicle weight decreases as described above, in addition to attitude control, by strengthening the prevention of lifting of the inner wheels, it is possible to eliminate unexpected deterioration of the steering characteristics and also prevent the loss of turning stability. do not have.

Furthermore, if the wheels that reduce the amount of load movement are drive wheels, loss of driving force can also be eliminated.

(Second Embodiment) Next, a second embodiment, which is an embodiment of the invention as claimed in claim (2), will be described based on FIG. 5 of the accompanying drawings.

This second embodiment has the same configuration as the first embodiment described above, and corrects the drive motor rotation angle θ9 as a command value in the manner shown in FIG. Here, the same reference numerals are used for the same components as in the first embodiment.

The controller 36 performs the process shown in FIG. 5 through the same steps as in FIG. On the coordinate axes in the figure, as the vehicle weight W increases, the saturation start point is moved to the side where the lateral acceleration α7 is higher. For example, when W=WL during stacking, when the lateral acceleration is 1α71-1αYl1 or more, the drive motor rotation angle 1θ, 1 is saturated, and its maximum value (saturation value)
]θ, the command value characteristic suppressed to l=lθMll is imported. However, if W=WH during heavy loading, the lateral acceleration is]
When αy 1−α7□ or more, a command value characteristic in which the rotation angle 1θ, 4 ] is suppressed to the maximum value 1θM21 (->16M11) is taken in.

The other configurations are the same as the first embodiment.

Therefore, according to the second embodiment, even in a turning state where the same lateral acceleration α occurs, as the vehicle weight W decreases, the maximum value of the drive motor rotation angle 1θ8 is lowered earlier, and the distance between the left and right wheels is reduced. The amount of load transfer also decreases. Therefore, even if the lateral acceleration becomes large, there is hardly any excessive load movement, and similarly to the first embodiment described above, lifting of the inner race is accurately prevented, and deterioration of steering characteristics is also eliminated.

(Third Embodiment) Next, a third embodiment, which is an embodiment of the invention set forth in claim (3), will be described based on FIG. 6 of the accompanying drawings.

This third embodiment has the same configuration as the first embodiment described above, and performs the processing shown in FIG. 6 instead of the processing shown in FIG. 3 described above. Here, the same reference numerals are used for the same components as in the first embodiment.

The process shown in FIG. 6 is the same as that shown in FIG. 3, with a new step (a) added. The controller 36 is a step ■
, ■ to calculate the vehicle weight W, step ■a
to move to. In this step ■a, it is determined whether the calculated vehicle weight W is less than or equal to a preset reference value W1, and W>
In the case of W, the same processing as in FIG. 3 after step (2) is performed, but in the case of W≦W1, the process returns to step (2).

Therefore, step ■ until vehicle weight W>W,
The process of ■ is repeated, and the hydraulic stabilizer 1 is actually
Wheel load movement control according to No. 4 is prohibited. Here, the process of step (a) in FIG. 6 corresponds to the control inhibiting means of the invention described in claim (3).

Other processing is the same as in the first embodiment.

Therefore, according to the third embodiment, when the vehicle weight is low, such as when the vehicle is heavily loaded, the exertion of roll rigidity is left to the shock absorber 1° of the suspension, and the hydraulic stabilizer 14 discontinues its control. , you can concentrate on eliminating excessive load transfer. Therefore, as in each of the embodiments described above, lifting of the inner ring is accurately prevented.

In addition to the hydraulic stabilizer related to the X piping as described above, the wheel load movement control device of each invention of the present application includes, for example, a damping force control device that inserts variable damping force shock absorbers in the left and right wheels and controls the damping force of the variable damping force shock absorbers. It may be implemented by a stabilizer device that utilizes the torsional rigidity of a torsion bar. 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. No. 295714) may be used as a wheel load movement control device to perform the above-mentioned correction process based on the vehicle weight.

Further, the turning state detection means in each invention of the present application may be configured to estimate and calculate lateral acceleration from the steering angle and vehicle speed as described above, or may directly detect inertial force with a lateral acceleration sensor.

Furthermore, as the weight detection/estimation means of each invention of the present application, in addition to using the wheel load sensor described above, it is also possible to estimate the vehicle weight from the stroke amount of the suspension at that time. On the other hand, the weight detected or estimated by this weight detection/estimation means is not limited to the weight of the entire vehicle including the live load; for example, it detects or estimates only the live load, and the detected or estimated value The load movement command value may be corrected accordingly.

〔Effect of the invention〕

As explained above, in the invention described in claims (1) and (2), as the vehicle weight or the loaded weight decreases, the command value characteristic regarding the left and right load movement is corrected so that the amount of load movement becomes smaller, for example, Set the command value by referring to this corrected command value characteristic corresponding to the turning condition (such as reducing the overall command value characteristic according to the weight, or advancing the saturation time of the command value characteristic for lateral acceleration, etc.) However, since the actuator is controlled according to this set command value, even when the vehicle weight is small, such as in a light-duty vehicle, the amount of load transfer between the left and right wheels exceeds the initial load on the inner wheel due to this control. This has the effect of being able to appropriately avoid the situation, thereby preventing deterioration of steering characteristics and destabilization of turning caused by unexpected lifting of the inner wheel during turning. Furthermore, when such wheels are drive wheels, loss of the drive wheels due to lifting of the inner wheels can also be eliminated.

Furthermore, in the invention described in claim (3), when the vehicle weight or the loaded weight becomes less than a predetermined value, load transfer control is stopped, so roll rigidity is exerted by other suspension devices. , lifting of the inner ring due to excessive load movement is almost certainly prevented.

[Brief explanation of drawings]

Figures 1 (a) and (b) are claims correspondence diagrams of the claimed invention;
4 to 4 are diagrams showing the first embodiment, FIG. 2 is a schematic configuration diagram, FIG. 3 is a schematic flowchart showing an example of processing in the controller, and FIG. 4 is a map of the rotation angle of the drive motor. FIG. 3 is a correction characteristic diagram stored in FIG. FIG. 5 is a diagram showing the second embodiment, and is a correction characteristic diagram stored in the drive motor rotation angle map. FIG. 6 is a diagram showing the third embodiment, and is an example of processing in the controller. It is a schematic flow chart showing. The main symbols in the diagram are 6...vehicle body, 8...suspension link, 14...hydraulic stabilizer, 2OL. 20R...hydraulic cylinder, 26A, 26B...first
Hydraulic piping, 28A, 28B...Second hydraulic piping, 3
0...Control cylinder, 32A, 32B...
3rd hydraulic piping, 34... electric motor, 36... controller, 38... vehicle speed sensor, 39... steering angle sensor, 40a-40d... wheel load sensor.

Claims (3)

[Claims] (1) In a wheel load transfer control device that includes an actuator whose roll stiffness can be changed between the vehicle body and the wheels and controls the amount of movement of the wheel load by changing the operating state of the actuator, a signal indicating a turning state is sent. A turning state detection means for detecting, a command value setting means for setting a command value for load movement by comparing the detection signal of the turning state detection means with a command value characteristic regarding load movement between left and right wheels, and this command value setting. actuator control means for controlling the operation of the actuator according to a command value set by the means; and a weight detection means for detecting or estimating the weight of the vehicle or the loaded weight.
A wheel characterized by being provided with an estimation means and a command value characteristic correction means for correcting the command value characteristic so that the smaller the detected value or estimated value of the weight detection/estimation means, the smaller the load movement amount. Load transfer control device. (2) The command value characteristic correction means is a means for advancing the saturation timing of the command value characteristic as the detected value or estimated value of the weight detection/estimation means becomes smaller. Wheel load transfer control device. (3) In a wheel load movement control device that includes an actuator whose roll stiffness can be changed between the vehicle body and the wheels, and controls the amount of movement of the wheel load by changing the operating state of the actuator, a signal indicating a turning state is sent. A turning state detection means for detecting, a command value setting means for setting a command value for load movement by comparing the detection signal of the turning state detection means with a command value characteristic regarding load movement between left and right wheels, and this command value setting. actuator control means for controlling the operation of the actuator according to a command value set by the means; and a weight detection means for detecting or estimating the weight of the vehicle or the loaded weight.
A wheel load transfer control characterized by comprising an estimation means and a control prohibition means for prohibiting the operation of the actuator control means when the detected value or estimated value of the weight detection/estimation means becomes a predetermined value or less. Device.