JPH02296513A - Active suspension - Google Patents

Abstract

PURPOSE:To make a sharing ratio of axle load easily calculable without using any load detector by calculating the longitudinal sharing ratio of the axle load in use of a command value operated for level controlling, and changing the longitudinal sharing ratio of roll control force contrary to this calculated sharing ratio. CONSTITUTION:In a controller 32 which inputs each output signal of stroke sensors 30FL-30RR outputting a voltage signal proportionate to relative height between each wheel and a car body, there is provided a level control part 50 which performs each level control separately in front and in the rear of a vehicle. Also there are provided a roll rigidity operation part 52 finding roll rigidity as total roll control force, and a roll rigidity longitudinal distributing operation part 54 distributing this roll rigidity separately in front and in the rear of the vehicle. In addition, there is provided a longitudinal distribution ratio decision part 56 which determinals a roll rigidity longitudinal distribution ratio to be done on the basis of the level controlled result, and a longitudinal sharing ratio of axle load is successively operated hereat, while a longitudinal distributing ratio of the roll control force over roll rigidity, antiroll moment or the like is changed contrary to a load sharing ratio by the operation part 54.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to an active suspension for a vehicle, and in particular has a fluid pressure cylinder interposed between each wheel and a vehicle body, and the cylinder pressure is This invention relates to an active suspension that controls vehicle height and vehicle body changes.

(Prior Art) Conventionally, as this type of active suspension, for example, the one described in JP-A No. 63-41225 (the title of the invention is “
``active suspension control device'') is known.

This conventional active suspension calculates the axle load sharing ratio between the front and rear of the vehicle based on the output values of actuators such as hydraulic cylinders installed at each wheel and between the wheels, and load detectors such as load cells installed at each wheel position. a calculation means for calculating, and an adjustment means for adjusting the front-rear distribution ratio of the load transfer amount between the left and right wheels of the vehicle during turning, according to the front-rear axle load distribution ratio calculated by the calculation means; The operating force of each actuator is controlled based on the adjusted front-rear distribution ratio. In other words, we focused on the fact that when the front-rear distribution ratio of left-right load transfer during turning changes, the steering characteristics change due to the non-linearity of tire characteristics. This is offset by controlling the front and rear distribution ratio of the amount of movement.

[Problems to be Solved by the Invention] However, in such conventional active suspensions, each wheel is provided with a t[detector such as a load cell,
The axle load sharing ratio between the front and rear of the vehicle can be calculated from (1) the detection results of the load detector when the vehicle is stopped, or (2) the detection results of the load detector during steady driving. Therefore, the load detector is expensive, which increases the manufacturing cost of the entire device, and as the number of sensors in the entire device increases, the system reliability increases accordingly.
■There was a problem that the quality decreased.

Furthermore, the method (1) above has the problem of not being able to deal with changes in the load sharing ratio due to fuel consumption during long-distance driving, and the method (2) does not allow for accurate calculation of the load sharing ratio. In order to cut the transient load vibration caused by manual force through the hydraulic cylinder, it is necessary to perform averaging processing using a low-pass filter or logical judgment processing for the vibration amplitude. There is also the problem that the calculation speed is forced to decrease due to the calculation load.

The present invention was made in view of such conventional problems, and it is possible to easily calculate the axle load sharing ratio between the front and rear of a vehicle without using an expensive load detector, thereby simplifying the configuration and The problem to be solved is to reduce manufacturing costs and to always be able to obtain constant vehicle steering characteristics.

[Means to solve the problem]

In order to solve the above problems, we have developed a hydraulic cylinder installed between the vehicle body and each wheel, a control valve that controls the operating pressure of each hydraulic cylinder according to a command value, and a control valve that controls the operating pressure of each hydraulic cylinder according to a command value. A vehicle height control means that calculates a command value for correcting a deviation from a target vehicle height value for each front and rear wheels, and a front and rear distribution ratio of a roll suppressing force that determines a command value for correcting the roll of the vehicle body and can change the command value to the command value. In an active suspension equipped with an attitude control means that calculates command values for the front and rear wheels by multiplying the
a front-rear load sharing ratio calculation means for calculating a front-rear axle load sharing ratio based on command values for adjusting the vehicle height of the front wheels and rear wheels calculated by the vehicle height control means; and a front-rear distribution ratio changing means for changing the front-rear distribution ratio of the roll suppressing force contrary to the distribution ratio.

[Effect]

In this invention, the front and rear axle load sharing ratio is calculated based on command values for front wheel side and rear wheel side vehicle height adjustment, which are calculated when the vehicle height control means controls the vehicle height value to the target value. The calculation means performs calculations sequentially. Therefore, the front-rear distribution ratio changing means changes the front-rear distribution ratio of roll restraining force such as roll stiffness and anti-roll moment in the posture control means, contrary to the load sharing ratio. For this reason, for example, if the load sharing ratio on the front wheels is larger than on the rear wheels, the roll suppression force distribution ratio on the front wheels will be smaller than that on the rear wheels, and the roll stiffness or anti-roll moment due to the attitude control means will be lower on the front wheels. is smaller than that of the rear wheels. Therefore, the shared load on the front wheels becomes larger than that on the rear wheels, and the steering characteristics change from strong understeer to oversteer. steering characteristics can be obtained.

At this time, a conventional load detector such as a load cell is no longer necessary in order to calculate the load sharing ratio before and after the load.

〔Example〕

Embodiments of the present invention will be described below based on the drawings.

(First Embodiment) FIGS. 1 to 4 are diagrams showing a first embodiment of the present invention.

In Figure 1, 10FL and 10FR are the front left and right wheels,
10RL and 10RR are rear left and right wheels, 12 is a wheel side member, 14 is a vehicle body side member, and 16 is an active type suspension, respectively.

Among these, the active suspension 16 is for each wheel 10FL.
- Hydraulic cylinders 18FL to 18RR and coil springs 19FL to 19RR as fluid pressure cylinders inserted in positions 10RR and 18FL to 18RR, respectively, and coil springs 19FL to 19RR, and
Pressure control valves 20FL to 201 as control valves that individually control the operating pressure of 18RR based on the command value I? R, a hydraulic pressure supply device 22 which is a hydraulic pressure source of this hydraulic system and includes a pump and a tank, this hydraulic pressure supply device 22 and a pressure control valve 2.
Accumulator 25.25 for accumulating pressure provided on the supply side between 0FL and 20RR, and each wheel 10FL-10R
Stroke sensors 30I'L~ provided at R positions respectively
30RR, detection signal H of this sensor 30FL to 30RR
It is equipped with a controller 32 that calculates a command value IFL-Il111 based on FL-Hun.

The coil springs 19FL to 19RR support the static load of the vehicle body, and their spring constants are lower than those of conventional mechanical suspensions (for example, 0.5 kgf/+nm) to improve ride comfort and ground contact. It has become. Furthermore, in order to improve riding comfort and ground contact, the conventional stabilizer for reducing roll is also omitted, and the inherent roll rigidity of the vehicle is lower than that of the conventional vehicle.

Each of the hydraulic cylinders L8FL to L8IIR has a cylinder tube 18a attached to the vehicle body side member 14, a piston rod 18b attached to the wheel side member 12, and a pressure chamber spaced from the piston 18c in the cylinder tube 18a. It is formed.

This pressure chamber communicates via a throttle valve 34 with an accumulator 36 for absorbing relatively high-frequency hydraulic vibrations.

Moreover, each of the pressure control valves 20FL to 20RR has a well-known structure (for example, Japanese Patent Laid-Open No.
2-295714), its supply port and return port are connected to the hydraulic pressure supply device 22 via piping 40° 42, and its output port is connected to the pressure chambers of the hydraulic cylinders 18FL to 18RR via piping 44. are connected to each. From the controller 32, each pressure control valve 20FL~
The command value I(
IFL~111R) are supplied respectively.

As shown in FIG. 2, each of the pressure control valves 20FL to 20RR supplies a pressure P proportional to the command value I from its output port to the hydraulic cylinders 18F to 18RR. In other words,
When the command value I is zero, a predetermined offset pressure P0 is output, and when the command value I increases in a positive or negative direction from this state, a proportional gain K is generated.

It outputs a pressure PC that increases or decreases with .

In addition, in FIG. 2, P□8 is the line pressure of the hydraulic pressure supply device 22.

On the other hand, the stroke sensors 30FL to 30RR are constituted by potentiometers in this embodiment, and are attached between the wheel side member 12 and the vehicle body side member 14 at each wheel position. Therefore, each of the sensors 30FL to 30RR generates a stroke signal HF, which is a voltage signal proportional to the relative distance between the vehicle body and the wheels at each wheel position, that is, the relative height.
L"' HRR is output to the controller 32.

As shown in FIG. 3, the controller 32 includes a vehicle height adjustment section 50 that adjusts the vehicle height separately for the front and rear of the vehicle, a roll stiffness calculation section 52 that calculates roll stiffness as a total roll suppression force, and a roll stiffness calculation section 52 that calculates the roll stiffness for each of the front and rear of the vehicle. A roll rigidity front and rear distribution calculation unit 54 that distributes rigidity, and a front and rear roll rigidity distribution ratio determination unit 56 that determines a roll rigidity front and rear distribution ratio based on the vehicle height adjustment result.
and an adder 5 provided separately for the pressure control valves 18FL to 18RR.
81"L to 58RR.

In this embodiment, stroke sensors 30FL to 30RR,
Vehicle height adjustment section 50. and adders 58FL to 58RR constitute vehicle height control means, and front left and right stroke sensors 30FL,
30FR, roll rigidity calculation section 52. Roll stiffness front and rear distribution calculation unit 54. and adders 58FL to 58RR constitute attitude control means.

The vehicle height adjustment unit 50 generates vehicle height adjustment command values V,
, V, front and rear average vehicle height calculation circuits 60F, 60R1 target vehicle height command circuits 62F, 62R.
5 adders 64F, 64R and integrators 66F, 66R.

The front average vehicle height calculation circuit 60F inputs the detection signals Hrt and H□ of the front stroke sensors 30FL and 30FR and calculates the front average vehicle height value (IPAV) as HFAV =
(HFL + HF11) / 2, and outputs a voltage signal corresponding to the average vehicle height value HFAv to the "+" terminal of the adder 64F. The target vehicle height value command circuit 62F is configured to output a voltage signal corresponding to a preset target vehicle height value Hor to the "=" terminal of the addition 2364F. Therefore, the adder 64 calculates 11r = HF
By calculating AV HOF, a voltage signal corresponding to the front vehicle height deviation hv is obtained, and the deviation is output to the integrator 66F. The integrator 66F has V,=A with respect to the input deviation signal hF.

- Perform an integral calculation of 5tlr-dt to obtain a front vehicle height command value (■), and add this to the adder 58FL on the front wheel side.

"-1" by outputting to the "-" terminal of 58FR
The vehicle height is adjusted in the opposite direction to the vehicle height change by multiplying by

Similarly, on the rear wheel side, the rear average vehicle height calculation circuit 60R receives the detection signals HllL+HIR from the rear stroke sensors 30RL and 30RR and calculates the rear average vehicle height value HMA.
V (=(HRL+H□)/2)], the adder 64R calculates the rear vehicle height deviation h++ (=HRAv HOR:
, -dtl are calculated and added to the adder 58RL on the rear wheel side.

It is designed to be output to the "-" terminal of 58RR.

Here, AF and AR are vehicle height adjustment gains that determine the speed of vehicle height adjustment. In addition, block 66 in FIG.
F, S in 66R represents the Laplace operator.

The roll stiffness calculation section 52 includes a roll angle calculation circuit 68 and a total roll stiffness calculation circuit 70. The roll angle calculation circuit 68 calculates the vehicle roll angle from the left-right difference in front vehicle height, and specifically, detects the detection signals H, L, H,
, and calculates φ=(■1□−HFL)/2, and outputs a voltage signal corresponding to the roll angle φ to the total roll stiffness calculation circuit 7°. This arithmetic circuit 70 is
The input roll angle signal φ is multiplied by a predetermined gain to obtain the total roll stiffness φ, and a voltage signal corresponding to φ is applied to the roll stiffness at the next stage, the roll stiffness longitudinal distribution calculation section 54.
Output to.

This roll stiffness front and rear distribution calculation section 54 includes a front distribution calculation circuit 72. It is configured with a rear distribution calculation circuit 74. The front distribution calculation circuit 72 multiplies the signal corresponding to the total roll stiffness φ by a roll stiffness front distribution ratio α (0≦α≦1) determined as described later to obtain the Cole stiffness that the front carries. A command signal R corresponding to the roll stiffness is output to the "+" terminal of the adder 58FL and the "-" terminal of the adder 58FR in opposite directions. The rear distribution calculation circuit 74 multiplies the total roll rigidity by φ by the rear distribution ratio of roll rigidity (l−α) to determine the roll rigidity that the rear carries, and sends a command signal R corresponding to this roll rigidity to the adder 58.
It is output to the "+" terminal of RL and the "-" terminal of 58RR, respectively.

Further, the front-rear distribution ratio determining section 56 is for the front.

It has rear reference value generation circuits 76F and 76R, front and rear axle load calculation circuits 78F and 78R1 load ratio calculation circuit 80, and roll stiffness front and rear distribution ratio calculation circuit 82.

Of these, the front reference value generation circuit 76F sends a signal corresponding to the static load WF0 shared by the front wheel side coil springs 19FL and 19FR to the front axle load calculation circuit 78.
The rear reference value generation circuit 16R outputs a signal corresponding to the static load WRo shared by the rear wheel side coil springs 19RL and 19RR to the rear axle load calculation circuit 78R.

The front load calculation circuit 80F calculates WF = based on the input reference signal WFo and vehicle height adjustment command signal ■1.
WFO+SF ・■.

=Wro+Sy ・AF ・S ((HrL+H
r++)/2-HOF) -dt -0
) to determine the axle load WF acting on the front wheels 10FL and 10FR. Similarly, rear load calculation circuit 8
0R is based on the input reference signal WRo and vehicle height adjustment command signal 8.
+ HRR)/2-HOR) ・dt
... Perform the calculation in (2), rear wheels 10RL, 10
Find the load WR acting on RR. Here, SF, SR
is on the front wheel side.

Rear wheel side hydraulic cylinder 18FL (18FR), 1
The pressure receiving area is 8RL (18RR).

In other words, the output V of the integrators 66F and 66R related to vehicle height adjustment
, ,V, are the front axle loads, respectively.

Since the value corresponds to the rear axle load, the above (102)
The operation of the expression is established. In equations (1) and (2), the first item is the static load shared by the coil spring, and
The spring constant is determined by the spring constant and the difference between the free length of the spring and the current length. When the vehicle height is at the target value (constant), the above-mentioned "difference between the free length of the spring and the current length" can be treated as a constant, so WFo and WRo can also be treated as constants. Note that in equations (1) and (2), the second item is the static load shared by the hydraulic cylinder.

Further, the load ratio calculation circuit 80 receives the output signal WF of the front and rear axle load calculation circuits 78F and 78R.

In this embodiment, by inputting Wl, the ratio of front side axle load sharing ratio Wr / (WF + WR) is calculated, and a signal corresponding to this ratio is output to the next stage front-rear distribution ratio calculation circuit 82. It is. In this embodiment, as shown in FIG. 4, the front/rear distribution ratio calculation circuit 82 calculates the distribution ratio 'WF/<wr +
When a signal corresponding to WR) J is input, a signal corresponding to a front side roll stiffness distribution ratio α (0≦α≦1) which is inversely proportional to this is input, and this signal is sent to the front/rear distribution calculation circuit 72F, It consists of a function generator that supplies 72R.

Of the above-mentioned configuration, in this embodiment, the front and rear reference value generation circuits 761"', 76R, the front and rear axle load calculation circuits 78F and 78R3, and the load ratio calculation circuit 80 constitute load front and rear distribution ratio calculation means. , front/rear distribution ratio calculation circuit 8
2 corresponds to the front/rear distribution ratio changing means.

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

Now, assume that the vehicle is in a standard load state, and the front and rear vehicle height values match the target vehicle height values H8v and R08. At this time, the front and rear deviation hr calculated by the controller 32 becomes HR=O, so the vehicle height adjustment command values V, , V, I=O, which are the outputs of the integrators 66F and 66R, are commanded. Value IFL~II
IN is only a component related to roll control. However, assuming that no roll occurs, the roll suppression command value R,,
Since R,=0, IFL" I**=O, and each pressure control valve 20FL~20RR has IFL~
An offset pressure P corresponding to IRR=O is output to each hydraulic cylinder 18FL to 18RR, and a force related to this pressure P0 and a coil spring 19FL to 19RI? The spring force supports the vehicle body at the target vehicle height.

However, suppose that the vehicle height value also deviates from the target value due to changes in the riding status of passengers, loading status of cargo, fuel consumption, etc., and changes in the vehicle attitude. This change is captured as a change in the stroke detection signal HFL"' HR11,
Average vehicle height value for front and rear HFAV + HRav
deviates from the value under standard load. Therefore, the deviation, h, corresponding to the change in vehicle height is obtained separately for the front and rear, and this deviation, h, is integrated, and the gain A,,A,
The vehicle height adjustment command values V, , Vll are calculated. Therefore, assuming that no roll occurs, the overall command values IFL, IFR = VF, IRL, [RR
= V*, and the control pressure P in the direction opposite to the vehicle height change.

is from the pressure control valves 20FL to 20RR to the hydraulic cylinder 18.
['L to 18RR, so the vehicle height values are controlled to gradually approach the target values H6F+HOR for the front and rear separately. This vehicle height adjustment has a deviation hF.

This is performed until h*=o. Since the vehicle height adjustment in this embodiment is substantially a PI operation, the vehicle height is adjusted at a speed proportional to the deviation, and no vehicle height offset occurs.

The command values V, VR calculated by the controller 32 in connection with this vehicle height adjustment are values corresponding to the front axle load and the rear axle load, respectively. Therefore, as mentioned above, the front desk is based on (1) and (2) above.

The rear axle loads Wr and WR are sequentially calculated, the front side load sharing ratio WF/(Wr + WR) is calculated, and the front side distribution ratio α of roll stiffness is always set in inverse proportion to this sharing ratio. The rear side distribution ratio at this time is (1-α). That is, as the load sharing ratio on the front side increases, the roll stiffness distribution ratio α on the front side becomes smaller in inverse proportion to this.

On the other hand, if the vehicle body rolls due to turning, etc., this roll condition is reflected in the front side stroke detection signal HFL.
, I (reflected in FR. At this time, the front average vehicle height value HFAV calculated from the stroke detection values HFL and HF* is almost the same as the target vehicle height value H8F, so the vehicle height adjustment has little to do with cylinder pressure control. do not.

Therefore, as described above, the controller 32 calculates the roll angle φ from the stroke detection value HFLI HFR.
φ is calculated as the total roll stiffness. This total roll stiffness is multiplied by the distribution ratios α and (1-α) set in real time as described above, respectively, to calculate the roll suppression command values RF and RR. Therefore, the command value 1r
t=Rr, I□=Ry, IRL=R11, I+u+=
Since the left and right phases of RR are reversed, the hydraulic cylinders 18FL and 18RL (or 18RL and 18RR) on the outer ring side
) increases, and the inner hydraulic cylinder 18RL,
18RR (or 18FL).

18RL) decreases, a biasing force is generated that resists the roll, and the roll is suppressed.

When suppressing this roll, if the vehicle height is set to 11 with the front side axle load sharing ratio being larger than the rear side, the load sharing ratio WF / (WF + WR) will be large, that is, the roll stiffness distribution ratio α will be small. Therefore, the roll stiffness on the front side is smaller than on the rear side. As a result, the amount of load movement on the front side is smaller than that on the rear side, and the sum of the cornering forces is larger, resulting in a greater grip force.
Steering characteristics can be adjusted to oversteer side. In other words, in the conventional case, when the load sharing ratio (Wr / CWr + WR) is large, that is, when the front is heavy, the vehicle tends to understeer strongly, but in this embodiment, the vehicle automatically shifts to the oversteer side in such a state. This can reliably prevent strong understeer. On the contrary, the load sharing ratio Wr/
(WF + WR) is small (the rear wheel load is large), that is, the roll stiffness distribution ratio α is large, and it is possible to reliably prevent oversteer that would otherwise occur. In this way, while maintaining good ride comfort and ground contact, almost constant steering characteristics can be obtained at all times without being affected by load loading conditions.

In particular, in this embodiment, there is no need to limit the setting of the roll stiffness distribution ratio α when stopped or steady running as in the past, and an accurate ratio α can be set at all times. Since it can also respond to changes in the load sharing ratio due to changes in consumption, the steering characteristics will no longer change midway through the journey and cause an unnatural feeling.

In addition, by using the vehicle height adjustment command value, there is no need to provide a filter for averaging the detected load values, which was seen in the conventional example, and the controller can be simplified, and the load cell, etc. Since a detector is not required, component costs can be reduced accordingly, and the reliability of the system can be increased by reducing the number of sensors.

Note that the calculation of the roll angle in the above embodiment is based on the front wheels 10
Although the attitude control means of the present invention is determined from the difference between the left and right strokes of FL and 10FR, the attitude control means of the present invention is not necessarily limited to this, and may be determined from the difference between the left and right strokes of the rear wheels, for example.

(Second example) Next, a second embodiment will be explained based on FIG.

Here, the same components as in the first embodiment are given the same reference numerals.

In the first embodiment, the roll situation was determined from the stroke detection signal, so the configuration allowed the vehicle body to roll to some extent, but in the second embodiment, a roll flat suspension is used that prevents such roll from occurring. The present invention is applied to.

The roll moment that causes the vehicle to roll is proportional to the vehicle's lateral acceleration (acceleration in the vehicle width direction) (vehicle weight,
(Determined by tread, roll center height 9 center of gravity height)
If the lateral acceleration can be measured, it is possible to generate an anti-roll moment (roll suppression force) using hydraulic pressure before a roll occurs, thereby realizing a roll land.

Focusing on this, in the second embodiment, a lateral acceleration sensor 90 mounted, for example, at the center of gravity of the vehicle body is newly added to the configuration of the first embodiment, and the lateral acceleration detection signal of this lateral acceleration sensor 90 is G is input to the controller 32 (
Here, the drawings corresponding to the above-mentioned FIG. 1 are omitted). Therefore, as shown in FIG. 5, the controller 32 is provided with a gain setter 92 that calculates a roll suppression command value by multiplying the input lateral acceleration detection signal G by a total anti-roll gain. The output of this gain setter 92 is individually outputted to front/rear distribution calculation circuits 94 and 96 having the same configuration as the first embodiment. Both circuits 94.
Reference numeral 96 constitutes an anti-roll moment front-rear distribution calculating section 98, which calculates "K8 ・αJ, 'KM ・
(l−α)” and the resultant signal MF,
Mt+ is sent to adder 58F and output to 58RR.

Here, the lateral acceleration sensor 90. Gain setter 92, anti-roll moment front and rear distribution calculation unit 98° and adder 5
8FL to 58RR constitute attitude control means. In addition, in this second embodiment, the front.

It is assumed that α input to the rear distribution calculation circuits 94 and 96 is read as the front-to-rear distribution ratio of the anti-roll moment.

Other configurations and operations are the same as in the first embodiment.

For this reason, for example, when lateral acceleration G occurs when turning, the attitude control command values MF and MR according to this lateral acceleration G are
is calculated, and an anti-roll moment opposing the roll moment is generated before the roll, depending on the left and right cylinder pressures. This allows the vehicle to remain in a flat position while the first
It becomes possible to control the steering characteristics in the same manner as in the embodiment.

In each of the above embodiments, vehicle height control is performed only for the average vehicle height of the front wheels and the average vehicle height of the rear wheels, but by adding vehicle height control in the roll direction to this, the three degrees of freedom of the vehicle posture (
Roll, pitch, lift) may all be controlled.

Further, the fluid pressure cylinder in the present invention may be a pneumatic cylinder, and the pressure control valve may be replaced with a control mechanism such as a pressure sensor and a servo valve.

〔Effect of the invention〕

As explained above, according to the present invention, the front and rear distribution ratio of the axle load is calculated using the command value calculated for vehicle height adjustment, and the front and rear distribution of the roll suppressing force is contrary to the calculated distribution ratio. By changing the ratio, it is possible to easily calculate the sharing ratio of the axle load, which changes from moment to moment, with relatively simple calculations, without using the conventionally installed load detectors such as load cells. By simplifying the configuration, manufacturing costs and weight are reduced, and
While the reliability of the system is improved, the front-to-back distribution of left-right load transfer can be controlled according to the front-to-back distribution ratio of the roll suppression force, and almost constant steering characteristics can be obtained regardless of the vehicle's driving condition. This has the effect of improving maneuverability.

[Brief explanation of drawings]

FIG. 1 is a schematic configuration diagram showing a first embodiment of the present invention, and FIG.
The figure is a graph showing the output pressure characteristics of the pressure control valve applicable to the present invention, Figure 3 is a block diagram showing the controller of the first embodiment, and Figure 4 is a graph showing the axle load sharing ratio applicable to the present invention. FIG. 5 is a characteristic diagram of the front-rear distribution ratio of roll stiffness, and is a schematic configuration diagram showing a second embodiment of the present invention. In the figure, 12 is a wheel side member, 14 is a vehicle body side member, 16 is an active suspension, 18FL to 18RR are hydraulic cylinders as fluid pressure cylinders, 20FL to 20RR are pressure control valves as control valves, 32 is a controller, 30FL
~30RR is a stroke sensor, 50 is a vehicle height control unit, 5
2 is a roll rigidity calculation unit, 54 is a roll rigidity front and rear distribution calculation unit, 56 is a front and rear distribution ratio determination unit, 58FL to 58RR are adders, 90 is a lateral acceleration sensor, 92 is a gain setter, 9
8 is an anti-roll moment front-rear distribution calculating section. Figure Figure Pressure Figure 4n 1t-(WF +WR)

Claims (1)

[Claims] (1) Fluid pressure cylinders interposed between the vehicle body and each wheel, control valves that control the operating pressure of each fluid pressure cylinder according to a command value, and a control valve that controls the actual vehicle height value and target vehicle height value. A vehicle height control means for calculating a command value for correcting the deviation for each front and rear wheels, and determining a command value for correcting the roll of the vehicle body,
An active suspension comprising an attitude control means for calculating command values for front and rear wheels by multiplying the command value by a changeable longitudinal distribution ratio of roll suppressing force, wherein the front and rear wheels are calculated by the vehicle height control means. A front and rear load distribution ratio calculation means for calculating the front and rear distribution ratio of the axle load based on a command value for adjusting the vehicle height on the wheel side; An active suspension characterized by comprising a front/rear distribution ratio changing means for changing the distribution ratio.