JPH01160721A - Electronically controlled suspension unit - Google Patents

Abstract

PURPOSE:To improve control response in a transient turning state in the suspension unit in the caption for changing a vehicle steering characteristic by compensating a position control signal so as to increase a rolling rigidity distribution on a front-wheel side in a transient turning state of a vehicle rather than in its steady turning state. CONSTITUTION:A turning state is detected by a turning state detection means M3 according to a vehicle running state, so as to be inputted to a rolling rigidity distribution control means M4. And the rolling rigidity distribution control means M4 judges a transient turning state based on a vehicle turning state which is inputted, and outputs a compensation signal to a position control means M2 so as to increase a rolling rigidity distribution on a front-wheel side in a transient turning state rather than in a steady turning state. A fluid is supplied/exhausted to/from each suspension M1 by the position control means M2 according to the compensation signal. In this structure, control response in a transient turning state can be improved.

Description

DETAILED DESCRIPTION OF THE INVENTION OBJECTS OF THE INVENTION [Industrial Field of Application] The present invention relates to an electronically controlled suspension device capable of controlling a suspension and thereby changing the steering characteristics of a vehicle.

[Prior Art] Various devices have been proposed in the past that change steering characteristics by controlling the suspension and changing the roll stiffness distribution during turning. As shown in FIG. 26, there is a non-linear relationship between the loads on the inner and outer wheels during turning and the cornering power. As shown by arrow a in the figure, the cornering power when the moving load between the inner and outer wheels is small is the sum of the value CP2I on the inner wheel side and the value CP2O on the outer wheel side. On the other hand, as shown by the arrow in the figure, the cornering power when the moving load between the inner and outer wheels is large is the sum of the straight CPII on the inner wheel side and the value CPIO on the outer wheel side. The magnitude relationship between the above two phases is as shown in the following equation.

2 CP

Further, the steering characteristics of the vehicle are determined based on the value of the following formula.

C"◆Lr-Cf number Lf=Z Here, Cr is the cornering power of the rear wheel, and L"
is the distance between the rear wheel axis and the vehicle center of gravity, Cf is the cornering power of the front wheels, and Lf is the distance between the front wheels and the vehicle center of gravity. , If the value of Z is negative, it is oversteer, if it is zero, it is neutral steer, and if it is positive, it is understeer. Therefore, when turning, controlling to reduce the moving load on the front wheel side will increase the cornering power on the front wheel side, resulting in oversteer characteristics, while controlling to increase the moving load on the front wheel side will cause the cornering power on the front wheel side to decrease. This is known to cause understeer characteristics.

For this reason, for example, the suspension can be configured to be controlled using a control command value that corrects the deviation between the target stroke of the unsprung area and the unsprung floor and the actually measured actual stroke based on the detected lateral acceleration value. And the amount of load movement in the left and right direction on the rear wheel side is arbitrarily set according to the lateral acceleration,
A device has been proposed that can arbitrarily change the steering characteristics of a vehicle when turning (Japanese Patent Application Laid-Open No. 61-1939).
08).

[Problems to be Solved by the Invention] However, in a conventional general vehicle, the stability factor Kh, the yawing resonance frequency fy, and the ratio of the yaw rate YR to the steering angle MA have the following relationships.

Kh= (Wf/Cf-Wr/Cr)/(L ◆G)
f♂l=L conflict (Cf number C"・ (1/V2+K
h)/(M◆I)) ””/2π YR/MA= (V/ (1+Kh-V”))
/(L-N) Here, L is the wheelbase, Wf and Wr are the ground loads of the front and rear wheels, M is the mass of the vehicle, ■ is the unidirectional moment of inertia, and ■ is the vehicle speed. , and N is the steering wheel gear ratio.

Furthermore, the dynamic characteristics of the vehicle's yaw rate are shown in Fig. 25.
There is a relationship like a graph in which the vertical axis shows the ratio of YR (yaw rate) to the steering angle MA and the phase difference between the steering angle and the yaw rate in terms of frequency, and the horizontal axis shows the steering speed.

In other words, in a general vehicle that experiences roll due to the front and rear suspensions as shown by the solid line in Figure 25, the yaw rate of the vehicle changes depending on the steering speed, and as the steering speed increases, a phase lag occurs and the yaw rate response changes. There have been problems in that the response of the yaw rate may be insufficient due to phase lag, and the steering/rudder response characteristics may not be sufficiently improved. Also, when turning, if the moving load on the front wheel side is reduced and set to oversteer, the cornering power C on the front wheel side will be reduced as described above.
As f increases, as is clear from the above equation, the stability factor Kh decreases, the ratio of yawley (YR) to steering angle MA increases, and the effectiveness of the rudder increases;
There is a problem in that the yawing resonance frequency fy, that is, the responsiveness of the yaw rate decreases.

SUMMARY OF THE INVENTION An object of the present invention is to solve the above-mentioned problems, and to provide an electronically controlled suspension device that improves steering response during a transient state of turning.

[Means for Solving the Problems] In order to achieve the above object, the present invention has the following configuration as a means for solving the problems. That is, as illustrated in FIG. 1, an attitude control means controls the attitude of the vehicle during turning by supplying and discharging fluid to the suspension M1 provided corresponding to the wheels of the vehicle, with a predetermined front and rear wheel roll stiffness distribution. In the electronically controlled suspension system having M2, a turning state detecting means M3 detects a turning state from the running state of the vehicle, and based on the turning state detected by the turning state detecting means M3, during a transient turning state, the turning state is lower than during a steady turning. Attitude control means M2 sends a correction signal to increase roll stiffness distribution on the front wheel side.
This is the configuration of an electronically controlled suspension system characterized by comprising: a roll stiffness distribution control means M4 outputting an output to a roll stiffness distribution control means M4;

[Operations] In the electronically controlled suspension system having the above configuration, the turning state detection means M3 detects the turning state from the running state of the vehicle, and the roll stiffness distribution control means M4 detects the turning state detected by the turning state detection means M3. Based on this, a correction signal is output to the attitude control means M2 to increase the roll stiffness distribution on the front wheel side during a turning transition state than during a steady turning state, and the attitude control means M2 supplies and discharges fluid to the suspension M1 according to the correction signal. do. Therefore, it is possible to improve the steering response during the turning transient state.

[Example code] Embodiments of the present invention will be described in detail below based on the drawings.

FIG. 2 is a schematic configuration diagram of an electronically controlled suspension device according to an embodiment of the present invention, and FIG. 3 is an air circuit diagram of the electronically controlled suspension device of this embodiment. This electronically controlled suspension device includes a suspension IFL on the left side of the front wheel, a suspension IFR5 on the right side of the front wheel, a suspension IRL on the left side of the rear wheel, and a suspension IRR on the right side of the rear wheel, which are each connected to an air circuit AC.
FL, IFR, IRL, and IRR each have a gas spring 2P.
L, 2PR, 2RL, 2RR and shock absorber 3F
L, 3FR, 3RL, and 3RR are provided.

These gas splashes 2FL, 2PR, 2RL, 2RR are the third
As shown in the figure, each main gas chamber 4PL.

4FR, 4RL, 4RR and auxiliary gas chamber 5FL, 5FR, 5
Equipped with RL and 5RR, main gas chamber 4PL.

Some of 4FR, 4RL, and 4RR have diaphragm 6FL,
Since the main gas chambers 6FR, 6RL, and 6RR are formed, the vehicle height can be adjusted by supplying and discharging air to the main gas chambers 4FL, 4PR, 4RL, and 4RR. Also, gas splash 2FL, 2PR.

2RL and 2RR are spring constant switching actuators 7FL
, 7FR, 7RL, and 7RR, the main gas chambers 4FL, 4PR, 4RL, and 4RR and the sub gas chamber 5FL,
By communicating/blocking 5FR, 5RL, and 5RR or switching the air flow rate, the spring constant can be changed to "low,""medium," and "high." In addition, shock absorbers 3FL, 3FR, 3RL, 3RR1, f damping force switching actuators 8FL, 8FR, 8RL, 8RR
The damping force can be changed to ``low'', ``medium'', and ``high'' by driving the oil flow rate through the orifice in the piston.

On the other hand, each gas spring 2FL is provided in the air circuit AC.

A compressor 10 driven by a motor 9 is provided as a source of compressed air to be supplied to 2PR, 2RL, and 2RR, and the discharge side of this compressor 10 is connected to an air dryer 14 and exhaust air through a check valve 12 that prevents backflow. Each is connected to a switching valve 16. The air dryer 14 is equipped with one silica gel filter to remove moisture from the compressed air. This air dryer 14 has a fixed crest 18
and a check valve 20 that prevents backflow.
Shutoffable supply switching valve 22 and connection switching valve 24
are connected to each other. The other side of this supply switching valve 22 is connected to a relief valve 25 set at a predetermined pressure, and is also connected to a high pressure reserve tank 2 on the front wheel side via a high pressure reserve switching valve 26 that can be communicated and shut off. It is also connected to a high-pressure reserve tank 32 on the rear wheel side via a high-pressure reserve switching valve 30 that can be communicated and shut off. These high pressure reserve tanks 28.3
2 includes pressure sensors 34, 36 that detect the air pressure in the high-pressure reserve tank 28, 32, and a relief valve 38, 36 that is set to a predetermined pressure. 40 are arranged respectively.

Furthermore, the other one of the supply switching valves 22 is connected to the main gas chamber 4FL via a leveling valve 42 that can cut off communication φ.
It is connected to the main gas chamber 4FR via the leveling valve 44, to the main gas chamber 4RL via the leveling valve 46, and to the main gas chamber 4RR via the leveling valve 48, respectively. Each main gas chamber 4FL, 4FR, 4RL, 4R
R is a pressure sensor 50 for detecting air pressure. 52.
54° and 56 are connected to each other.

In addition, the main gas chamber 4FL on the left side of the front wheel is connected to the low pressure reserve tank 62 on the front wheel side through a discharge valve 58 that can be communicated with and shut off, and the main gas chamber 4PR on the right side of the front wheel is connected to a low pressure reserve tank 62 on the front wheel side through a similar discharge valve 60. has been done. Furthermore, the main gas chamber 4RL on the left side of the rear wheel is connected to the low pressure reserve tank 6 on the rear wheel side through a discharge valve 64 that can be communicated with and shut off, and the main gas chamber 4RR on the right side of the rear wheel is connected to the low pressure reserve tank 6 on the rear wheel side through a similar discharge valve 66. Each day is connected. On the other hand, the low pressure reserve tank 62 on the front wheel side and the low pressure reserve tank 68 on the rear wheel side are connected so as to be able to communicate at all times.・These low pressure reserve tanks 62 and 68 are equipped with pressure sensors 70. 72 are connected to each other, and a relief valve 74 set at a predetermined pressure is connected to the low pressure reserve tank 62 on the front wheel side.

Both of these low pressure reserve tanks 62 and 68 are connected to the other of the connection switching valves 24, and are connected to the compressor 10 via a suction switching valve 76 that can be communicated and shut off.
connected to the suction side of the

Further, a check valve 7 is connected to the suction side of the compressor 10 so that the air can be sucked in.

Without providing this check valve 78, the air circuit A
It is also possible to configure C as a completely closed circuit and to introduce air or other gas, such as nitrogen gas, into the air circuit AC.

The exhaust switching valve 16, the supply switching valve 22, the connection switching valve 24, the high pressure reserve switching valve 26.
30. Leveling valve 42. 44°46, 4B,
Discharge valve 5B, 60°64, 66,
In this embodiment, the suction switching valve 76 uses a normally closed type.

In this air circuit AC, a high pressure reserve tank 28.32 and a low pressure reserve tank 62.6 are provided on the front wheel side and the rear wheel side, respectively.
Although 8 is provided, one high pressure reserve tank and - number of low pressure reserve tanks may be provided in common on the front wheel side and the rear wheel side.

Furthermore, as shown in Figure 2, the distance between the left front wheel and the vehicle body,
That is, a vehicle height sensor 80 detects the left front vehicle height, a vehicle height sensor 82 similarly detects the right front vehicle height, a vehicle height sensor 84 that detects the left rear vehicle height, and a vehicle height sensor 84 that detects the right rear vehicle height. A vehicle height sensor 86 is provided respectively. Each of the vehicle height sensors 80°82, 84, 86 outputs a signal corresponding to a positive vehicle height difference when the vehicle height is higher than a predetermined reference vehicle height, and a negative signal when the vehicle height is lower than that. Outputs a signal according to the height difference. On the other hand, the vehicle speed is determined from a well-known steering angle sensor 90 that detects the steering angle of the steering wheel 88, a well-known acceleration sensor 92 that detects the lateral and longitudinal acceleration of the vehicle body, and the rotational speed of the output shaft of a transmission (not shown). A vehicle speed sensor 93 to detect, a door switch 94 provided for each door of the vehicle to detect the closed state of the door, a neutral switch 95 to detect that the shift state of the transmission is neutral, and an internal combustion engine (not shown). The throttle opening sensor 96 detects the opening of a throttle valve that regulates the intake air amount. In addition, by manual operation,
The vehicle is equipped with a vehicle height high switch 97 and a vehicle height low switch 98 for instructing the vehicle height.

Next, the electrical system of this embodiment will be explained using the block diagram shown in FIG. Each suspension IFL, IFR
, IRL, and IRR are driven and controlled by the electronic control circuit 100 to control the attitude of the vehicle. As shown in FIG. 4, this electronic control circuit 100 includes a well-known CPU 102. R
OM 104° RAM 106 is configured as the center of the logic operation circuit, and input/output circuits perform external and human output, here, an actuator drive circuit 10, a valve drive circuit 110,
A sensor input circuit 112, a level input circuit 114, etc. are connected to each other via a common bus 116.

The CPU 102 uses the pressure sensor 34. 3B, 50
゜52, 54. 56. 70. 72, vehicle height sensor 80°82, 84. 86, the signals from the steering angle sensor 90, acceleration sensor 92, vehicle speed sensor 93, and throttle opening sensor 96 are sent via the sensor input circuit 112 to the door switch 94, neutral switch 95, throttle opening sensor 96, and vehicle height high switch. inch 97 and vehicle height U-
The signal from the 9th switch is input via the level input circuit 114. On the other hand, these signals, ROM1'04,
Based on the data stored in the RAM 106, the CPU 102 controls the compressor motor 9 and the spring constant switching actuators 7FL and 7FR via the actuator drive circuit 10.
, 7RL, 7RR and damping force switching actuator 8F
A drive signal for driving L, 8FR, 8RL, and 8RR is output, and the exhaust switching valve 16 is output via the valve drive circuit 110.
, supply switching valve 22, connection switching valve 24, high pressure reserve switching valve 26.30, leveling valve 42.
44. 46°48, discharge valve 5B,
60. 64°66, outputs a drive signal to the intake switching valve 76, and outputs a drive signal to each suspension IFL, IFR, IRL, I
It controls RR.

The ROM 104 stores maps shown in FIGS. 10 to 22, which will be described later.

Next, the processing performed in the above-mentioned electronic control circuit 100 will be explained with reference to flowcharts shown in FIGS. 5 to 9.

FIG. 5 is a general flowchart showing an example of air suspension control according to the present invention, and FIGS. 6 to 9 are flowcharts showing detailed processing thereof.

The process shown in FIG. 5 is repeatedly executed at predetermined intervals.

First, when the process is started, it is determined in step 103 whether or not it is the first process after the power is turned on. If it is the first process, in step 105 various flags and variables are set for the amount of paddy. Next, in step 110, the output signals of the various sensors described above are read.

Next, at Stembu 200, the suspension IFI7 when the vehicle rolls. IFR, IRL, IRH gas spring 2PL
, 2FR, 2RL, and 2RR, feedforward control is executed. This feedforward control is applied to the vehicle by steering.
The predicted acceleration GRLM, which is the acceleration in the direction perpendicular to the vehicle traveling direction, that is, the acceleration in the lateral direction, is calculated, and the gas springs 2FL, 2PR, 2R are activated according to the predicted acceleration GRL.
L.

This is a control that attempts to prevent rolls or adjust to a predetermined inclination by adjusting the air pressure of the 2RR.

Next, in step 400, feedback control is executed among similar supply/discharge controls. This feedback control uses gas springs 2FL, 2PR, and 2 to stabilize the vehicle attitude when the acceleration of the vehicle is relatively stable.
This control attempts to adjust the air pressure of RL and 2RR.

Next, in the steering wheel 500, a corrected total pressure calculation for each wheel is executed, and the sum of the pressure correction amounts obtained by the feedforward control and feedback control is determined as the corrected total pressure. Calculate the roll stiffness distribution target value FM, and calculate the roll stiffness distribution target value FM.
This is a control that corrects the corrected total pressure according to.

Next, in step 800, high pressure reserve switching valves 26, 30, leveling valves 42, . 44. 46°48, and discharge valve 5B, 60. Valve control is performed to open and close necessary valves among 64 and 66.

The above feedforward control, feedback control section,
Details of corrected total pressure calculation for each wheel and valve control will be explained.

FIG. 6 represents a flowchart of feedforward control. First, in step 210, filtering processing is performed on each signal. In other words, the data read this time is
(n), the value after the previous filtering is Y (n-1
), and the filtering constant is If (=1 to 256), the output Y (n) due to filtering is expressed by the following equation.

This process is a process for canceling out noise in detected data and averaging fluctuations in data at a frequency higher than a predetermined value.

Next, a series of determination processes are performed to detect the state of the vehicle attitude change factor. That is, in step 220, the door switch 94 of the vehicle determines whether all the doors are closed, and in step 230, the neutral switch 95 determines whether the transmission is in the neutral state. In step 240, the throttle opening sensor 96 determines whether the throttle valve is fully open, and in step 250, among the suspension control valves, especially the high pressure reserve switching valves 26, 30,
Leveling valve 42. 44. 46. 4B, discharge valves 5B, 60, 64.66 determine whether suspension vehicle height control is being executed;
At step 260, the vehicle speed sensor 93 determines whether the vehicle speed is lower than a predetermined vehicle speed vO. Among these steps, steps 220, 230, 240, and 260 are the factors that change the attitude of the vehicle (door opening/closing indicating when a passenger gets on or off, transmission shift indicating the state of transmission of driving force to the tires, and internal combustion indicating the driving force itself). This is a step of detecting the state of the intake air amount to the engine and the vehicle speed indicating the running state, and step 250
is a step of detecting that gas is not being supplied or discharged to adjust the air pressure of the gas springs 2FL, 2FR, 2RL, and 2RR.

If all of these conditions are affirmatively determined, the attitude of the vehicle is stable and the gas is splashed 2PL.

Since no large pressure fluctuations have occurred in 2FR, 2RL, and 2RR, and it can be predicted that the atmospheric pressure is stable, in step 270, each pressure sensor 50 at that time is
．． 52.54.56 to each reference pressure P FLA, P
RAM as FRA, P RLA, P RRA
106. This pressure value is a value filtered with a lower frequency than the filtering in step 210, for example, with a 5Hz low-pass filter.
Filtering constant I used in step 210 above
f is set.

If even one of the above steps 220 to 260 is negative, step 270 is not executed, and each reference pressure P FLA, P FRA, P RLA ,
PRRA is not newly set. That is, as long as the conditions are met, each reference pressure P FLA, P F
RA, P RLA.

PRRA is updated.

Then after processing step 270 or step 220
~ If even one of the steps is negative in step 260, the estimated lateral acceleration dRL of the vehicle is calculated from the vehicle speed V and the steering angle θ based on the map shown in FIG. 10 in step 280. It will be done. The graph corresponding to the map in FIG. 10 shows only the cases of two different predetermined accelerations with two broken lines, and since the other relationships are the same, their description is omitted.

Of course, other acceleration values may be obtained by interpolation calculation.

Next, in step 290, the estimated lateral jerk 16RL of the vehicle is determined from the vehicle speed V and the steering angular velocity υ, which is the differential value of the steering angle θ, based on the map shown in FIG. Note that the steering angle υ may be a difference value of the steering angle θ within a predetermined period. The graph corresponding to the graph in FIG. 11 shows only the cases of eight different steering angular velocities O with eight broken lines, and the values between them are determined by interpolation calculation.

Next, in step 300, predicted acceleration GRLM is calculated by linear combination of the following formula.

GRLM = m ◆6RL+ h φ16RL where,
m and h are constants and have values appropriately determined through experiments or the like in order to predict the roll.

Next, in step 310, each suspension IFL,
IFR, IRL, IRR gas splash 2FL, 2FR, 2
Each target pressure difference ΔPFLM, ΔPFRM of RL and 2RR,
ΔPRLM and ΔP RRM are calculated. That is, if the horizontal axis is the predicted acceleration GRLM [G] and the vertical axis is the target pressure difference [kgf/cm2], then each target pressure difference ΔPFL
M, ∆PFRM, ∆PRLM, and ∆PRRM have a relationship as shown in the figure, and are expressed as in the following equation.

ΔPFLM= a ◆ GRLMΔPFRM
−-a ◆GRLM ΔPRLM= b Order GRLMΔP RRM=
-b◆GRLM Here, a and b are coefficients for correcting variations in suspension characteristics, and are expressed as in the following equation.

Here, W is the amount of sprung mass, h is the height of the center of gravity, tf is the front tread, tr is the rear tread, rf is the front arm ratio, rr is the rear arm ratio, Af is the front pressure receiving area, and A
r is the rear pressure receiving area, is the wheel base, and Lr is the distance between the rear wheel and the center of gravity. Also, Kf is (L/Lr)>K″f
5. When Kf=1.0, with an arbitrary value set in the range of ≧1.0, while respecting the front shared load increment. The shared load on the front will be 50%. By arbitrarily setting this Kf, it is possible to arbitrarily set the steering characteristics of the vehicle.

However, in order to avoid repeating minute adjustments due to calculations (surface runout, detection errors, noise, etc.), -1≦G
If RLM≦i, ΔPFLM=ΔP FRM=ΔP
RLM=ΔPRRM=O, and a dead zone is provided. Moreover, if the values of coefficients a and b are selected appropriately, each suspension IFL, IFR, and IRL can be adjusted.

IRR and high pressure reserve switching valve 26. 30. Leveling valve 42. 44. 46. Since the target pressure difference can be set according to the characteristics of the discharge valve 4B, the discharge valve 5B, and the discharge valve 60, 64, and 66, it is possible to eliminate functional errors between devices.

Next, in step 320, each target pressure PFLM.

P FRM, P RLM, and P RRM are calculated as shown below.

PFLM=ΔP FLM + P FLAPFRM=Δ
P FRM + P FRAPRLM=ΔP RLM
+P RLAPRRM=ΔPRRM+PRRA This determines each pressure to be controlled.

Next, in step 330, each pressure deviation eFL, eFR,
e RL and e RR are calculated as shown below.

eFL = PFLM - PFL eFR = PFRM - PFR eRL = PRLH - PRL e RR = PRRM - PRR where PFL, PFR, PRL, PRR are:
These are values obtained by filtering the outputs of the pressure sensors 50, 52, 54, and 56 provided in the main gas chambers 4FL, 4PR, 4RL, and 4RR of each suspension IFL, IFR, IRL, and IRR.

Next, in step 340, in order to convert each pressure deviation into a control manipulated variable, 1 is applied to each feedforward cane.
Based on the map shown by the dotted line in Fig. 13, the predicted acceleration G
It is determined according to the difference between RLM and actual lateral acceleration GRL. l When GRLM-GRLI is q or less, 1=0,
Above Q, kl = T, and in between, l GRLM
-GRL The relationship is such that it increases in accordance with an increase in l. However, k2°k3 represents the gain of feedback control, which will be described later. That is, the predicted lateral acceleration GR
This indicates that the greater the difference between LM and the current lateral acceleration GRL, the greater the contribution rate of feedforward control to the actual control amount.

Next, in step 350, using 1 as the gain and each pressure deviation e FL, e FR, e RL, e RR, each suspension IFL, IF
R.

Feedforward pressure fjkcI to IRL and IRR
FL, c IFR, c IRL, and c IRR are calculated.

c IFL = kl ◆eFL c IFR = kl No. eFR clRL = kl◆eRL clRR = kl Conflict eRR In this way, feedforward calculation processing is performed, and feedforward pressure m c IFL, c IF
R, c IRL and c IRR are calculated.

Next, the feedback calculation process shown in FIG. 7 is performed. First, in step 410, each suspension IFL
, the output values XFL, XFR, of the vehicle height sensors 80.82, 84.86 installed in IFR, IRL, and IRR,
According to XRL, XRRtZ, each displacement,
That is, the vertical displacement amount XH, the pitch displacement amount XP, the roll displacement amount XR5, and the torsional displacement amount X of the vehicle body are calculated.

XH= (XFR+XFL) + (XRR+XRL)
XP = (XFR+XFL) −(XRR+XRL)
XR= (XFR-XFL) + (XRR-XRL)X
W = (XFR-XFL) - (XRR-XRL) Here, XFR is the vehicle height on the right side of the front wheel, XFL is the vehicle height on the left side of the front wheel, XRR is the vehicle height on the right side of the rear wheel, and XRL is the vehicle height on the left side of the rear wheel. It represents the vehicle height.

Next, in step 420, each of the above displacement amounts XI.

Based on XP, XR, and XW, each mode deviation eH, eP, eR, and e- is calculated as shown in the following formula.

elf =XHM-XH eP =XPM-XP eR=XRM-XR eW=XWM-1 Here, X IIMc is the target vertical displacement amount, and
Vehicle speed V and vehicle height high switch 9 based on the map shown in the figure.
7 or the mode selected with the vehicle height low switch 9th (
H-AUTO or N-AUTO).

XPM is a target pitch displacement amount, which is determined from the actual acceleration GFR in the longitudinal direction of the vehicle detected by the acceleration sensor 92 based on the map shown in FIG. XRM is the target roll displacement amount, which is determined from the actual acceleration GRL in the vehicle lateral direction based on the map shown in FIG. X
WM is a target torsional displacement amount and is normally zero.

Next, in step 430, each of the above displacement amounts XH.

Based on the differential value statement H9 statement P9 statement R9 statement W of XP, XR, HJ, each mode speed deviation ^H, eP, e
R and ^W are calculated. Incidentally, the sentence H9 sentence P9 sentence R1 sentence W may be XH, XP, XR, XIvJ (D difference value for a predetermined period).

^■ = Bun HM - Bun H White P = Bundan - Bun P eR = Bun RM - Bun R ^W = Bun Kano - Bun W Here, Bun 11M is the target vertical displacement speed amount, which is normally zero. . Sentence PM is the target pitch speed displacement amount, and the first
Vessel speed d in the longitudinal direction of the vehicle based on the map shown in Figure 7
Determined from FR. Sentence RM is the target roll displacement speed amount, and is determined from the ship speed 6RL in the lateral direction of the vehicle based on the map shown in FIG. Sentence WM is the target torsional displacement velocity amount, which is normally zero.

Next, in step 440, 2H is applied to each feedback gain in order to convert each deviation into a control manipulated variable.

k2P, 2R, 2W (generally referred to as k2), and 3H, k3P, k3R, k3W (generally referred to as k3).

) is based on the map shown by the solid line in Fig. 13 mentioned above,
It is determined according to the difference between the predicted acceleration GRLM and the actual lateral acceleration GRL. l GRLM - GRt, if l is greater than or equal to Q, then 2, and 3=T; if l is greater than or equal to Q, then 2. 3=t
In between, the relationship is such that l GRLM - GRL decreases as l increases. That is, if the difference between the predicted acceleration GRLM and the current lateral acceleration GRL is small, the feed pump control falls to the actual control amount.
This shows that the contribution rate of

Next, in step 450, each mode deviation eH9eP,
eR, eW and each mode speed deviation white H1^P.

From ^R and eW, each feedback amount D is calculated as shown in the formula below.
H, DP, DR, and DW are calculated.

2L to DH= 3L to eH+ 4HDP to eH+
= to 2P◆eP + to 3P◆eP + to 4PDR= to 2
R order eR+ k3R◆e'F< 4RDW to +
: k2W ◆ eW + k3W conflict t=≦-J
+4W However, k4tl, k4P, k4R,
k4W is a predetermined constant.

Next, in step 460, each of the above feedback amounts DH
, DP, DR, and DW, each suspension IFL, 11"R, and IRL is determined by the following formula.

IRR feedback amount DFL, DFR, DRL
, DRR is calculated t times.

DFL= 1/4(k OH◆D H+2 k OPφLf◆DP
-k 0R-D R-k 01.1- D W) DFR
= 1/4 (k 011・DH+2kOP・1.f-DP+
kOR-DR+kOW-DW) DRL= 1/4 (kOH◆Dll-2kOP◆(1-Lf)◆D
P-k 0R4D R+ k 0W-D W) DRR= 1/4 (kOH◆DH-2kOP◆(1-Lf)・DP
+kOR"DR-kosu・DIJ) Here, kOH, kOP, kOR, kori represent predetermined coefficients, and Lf represents a distribution coefficient between the front and rear wheels taking into account the position of the vehicle center of gravity within the wheel base.

Next, in step 470, the feedback amount DE
Based on L, DFR, DRL, and DRR, each feedback pressure amount c2FL, c2FR is calculated using the following formula.
, c2RL, c2RR are calculated.

c2FL=PFL◆a2FL◆DFL c 2FR= P FR◆a2FR◆D FRc 2R
L-P RL ◆ a2RL Order DRLc 2RR-P
RR◆a2RR◆D RR where, PFL, PF
R, PRL, PRR are each suspension IFL,
IFR, IRL, IRR main gas chamber 4PL, 4FR
, 4RL, pressure sensors 50, 52 provided in 4RR,
This is the value obtained by filtering the output of 54.56. a
2FL, a2FR, a2RL, and a2RR are predetermined coefficients.

In this way, the feedback calculation process is performed, and the feedback pressure amounts c 2FL, c 2FR, c 2
R.L.

c 2RR is calculated.

Next, the pressure calculation process shown in FIG. 8 is performed.

First, the step 510 calculates the feedforward pressure amounts c IFL, c IFR, c IRL, c calculated in the feedforward calculation process as shown in the following formula.
IRR and the feedback pressure amount c 2FL, c 2FR, c 2 calculated by the above feedback calculation process
Corrected total pressure amount c FL from the sum of RL, c 2RR,
c FR, c RL, and c RR are calculated.

c Fl, = c IFL+ c 2FLc FR=
c IFR+ c 2FRc RL= c IRL+
c2RLcRR=cLRR+c2RRNext, in step 520, the roll stiffness distribution target value F is determined according to the estimated lateral jerk RL as a turning transient state calculated by executing the processing in step 290, as shown in the following formula.
M is calculated.

FM=KRH·1ii15RLl Here, KR)l is a front wheel side roll distribution force setting coefficient, and is a predetermined coefficient determined in advance through experiments or the like. Further, the roll stiffness distribution target value FM may be obtained from a map as shown in FIG. 19. The dot-dash line in this map is the same as when the above equation is shown as a map, and by using a map as shown in solid lines or broken lines, it is possible to It is also possible to set the optimum characteristic such as FM=0 or FM"=-50 depending on the characteristics of the car or the vehicle specifications.

Further, the calculation of the roll stiffness distribution target value FM is not limited to the estimated lateral jerk 6RL, but may be calculated using, for example, the differential value of the actual acceleration GRL detected by the acceleration sensor 92 or the actual lateral jerk GRL which is the difference value within a predetermined period. It may also be calculated based on the rate of change in the yaw rate, as long as it represents a turning transient state from a straight running state to a turning state where there is no change in lateral acceleration etc. due to steering.

Subsequently, in step 530, each suspension IFL
, IFR, IRL, IRR main gas chamber 4FL, 4FR,
Pressure sensors 50, 52.5 provided on 4RL and 4RR
4.56 output, filtering constant If =256
The values detected in, that is, the instantaneous values PFL, PFR, PR
According to L and PRR, suspension forces FFL, FFR, FRL, and FRR for each wheel are calculated as shown in the following formula.

FFL=Af◆1f φPFL FFR=Af◆lf ◆PFR FRL=Ar・1" ◆PRL FRR=A"φlr ◆PRR Here, A" is the pressure receiving area of front wheel suspension IFL, IFH gas splash 2PL, 2FR , Ar is the rear wheel suspension IRL, IRR gas spring 2RL, 2RR
1f represents the arm ratio at the position where the gas spring is installed on the front wheel side, and l' represents the arm ratio at the position where the gas spring is installed on the rear wheel side.

Next, in step 540, the absolute value of the difference between the suspension force FFL on the left side of the front wheel and the suspension force FFR on the right side of the front wheel is determined to be a predetermined torsional force guard value ΔF.
It is determined in step 550 whether the absolute value of the difference between the suspension force FRL on the left side of the rear wheel and the suspension force FRR on the right side of the rear wheel is greater than or equal to a predetermined torsion guard value ΔF. It is determined whether or not. In other words, if both of these conditions are affirmatively determined, the front wheel side of the vehicle body is twisted by the torsion force guard value ΔF or more, and the rear wheel side is also twisted by the torsion force guard value ΔF or more. It is a state that is occurring.

If an affirmative determination is made in both steps above, step 560
At step 570, it is determined whether the difference between the suspension force FFL on the left side of the front wheel and the suspension force FFR on the right side of the front wheel is greater than zero, and at step 570, the suspension force FRL on the left side of the rear wheel and the suspension force FRR on the right side of the rear wheel are determined. It is determined whether the difference between the two is greater than zero. If an affirmative determination is made in both steps, in other words, the suspension forces FFL and FRL of the left wheel are large for both the front and rear wheels, so it is determined that the vehicle is turning to the right with the left wheel being the outer wheel of the turn. Then, at step 580, the roll stiffness distribution target value FM is substituted into the stiffness distribution value FO.

Further, if a negative determination is made in step 570, that is,
If the front wheel side is receiving torsional force in opposite directions due to the left suspension force FFL and the rear wheel side is receiving torsional force in opposite directions due to the right suspension force FRR, in step 590,
Zero is assigned to the rigidity distribution value FO, and control for removing torsional force is given priority over changing the roll rigidity distribution.

On the other hand, if a negative determination is made in step 560, it is determined in step 600 whether or not the difference between the suspension force FRL on the left side of the rear wheel and the suspension force FRR on the right side of the rear wheel is smaller than zero. If an affirmative determination is made, that is, the right suspension force FFR, FR is applied to both the front wheel side and the rear wheel side.
Since R is large, it is determined that the vehicle is turning to the left with the right wheel on the outer wheel side of the turn, and in step 610, the negative roll stiffness distribution target (i-FM) is substituted for the stiffness distribution value FO.

Further, if a negative determination is made in step 600, that is,
If the front wheel side is receiving torsional force from the right suspension force FFR and the rear wheel side from the left suspension force FRL in opposite directions, then in step 590,
Zero is assigned to the rigidity distribution value FO, and control for removing torsional force is given priority over changing the roll rigidity distribution.

Furthermore, if a negative determination is made in step 550, that is, a torsion equal to or greater than the torsion force guard value ΔF has occurred only on the front wheel side, and no large torsional force has occurred on the rear wheel side. At 620, it is determined whether the difference between the suspension force FFL on the left side of the front wheel and the suspension force PFR on the right side of the front wheel is greater than zero. If a positive judgment is made,
In other words, since the suspension force FFL on the left side of the front wheel is large,
It is determined that the vehicle is turning to the right with the left wheel on the outer wheel side of the turn, and at step 630, the roll stiffness distribution target value FM is substituted into the stiffness distribution value FO. On the other hand, if a negative determination is made in step 620, that is, since the suspension force FFR on the right side of the front wheel is large, it is determined that the vehicle is turning to the left with the right wheel becoming the outer wheel of the turn, and step 6
At step 40, the negative roll stiffness distribution target value -FM is substituted into the stiffness distribution value FO.

Furthermore, if a negative determination is made in step 540, that is,
If a large torsional force is not generated on the front wheel side, but a torsional force is generated on the rear wheel side, at step 650, as in step 550, the suspension force FRL on the left side of the rear wheel is determined.
It is determined whether the absolute value of the difference between the suspension force FRR on the right side of the rear wheel and the suspension force FRR on the right side of the rear wheel is greater than or equal to a predetermined torsional force guard (direction ΔF).Here, if an affirmative determination is made,
At step 6GO, suspension force FR on the left side of the rear wheel
It is determined whether the difference between L and the suspension force FRR on the right side of the rear wheel is greater than zero. If a positive judgment is made, that is,
Since the suspension force FRL on the left side of the rear wheel is large, it is determined that the vehicle is turning to the right with the left wheel being the outer wheel of the turn, and in step 670, the roll stiffness distribution target value FM is substituted for the stiffness distribution value FO. do.

On the other hand, if a negative determination is made in step 660, that is,
Since the suspension force FRR on the right side of the rear wheel is large, it is determined that the vehicle is turning to the left with the right wheel being on the outer wheel side of the turn, and in step 680, the above-mentioned negative roll stiffness distribution target value -F is set to the stiffness distribution value FO. Substitute gate.

Furthermore, if a negative determination is made in step 650, that is,
Suspension force of each wheel FFL, FFR, FRL.

When the FRR is approximately equal and the vehicle is running straight, etc.,
At step 690, zero is assigned to the stiffness distribution value FO.

Next, in step 700, the suspension force FF
The vehicle body torsion force FW is calculated according to L, FFR, FRL, and FRR as shown in the formula below.

FW=(FFL-FFR)-(FRL-FRR) Note that it is also possible to use the torsion distribution ratio calculated as in the following formula instead of the torsion force FW.

CFl, l= (FFL-FFL) /((FFL-
FFR) + (FRL-FRR)) Subsequently, in step 710, the corrected total pressure amounts CFL, cFR, cRL,
cRR5 Rigidity distribution value FO5 Depending on the vehicle body torsional force FW, each torsion correction control pressure amount DFL, DFR, DRL,
DRR is calculated as shown below.

DFL= cFL+KIJF (FO-FW)DRL=
CRL-KWF< (FO-FW)DFR= cF
R-KWF (FO-FW) DRR= cRR+KWR
(FO-FW) Here, KWF and KWR are predetermined coefficients for converting force into pressure, which are determined in advance through experiments or the like.

That is, when it is determined that the vehicle is turning to the right, the roll stiffness distribution target value FM is substituted into the stiffness distribution value FO, and the front wheel suspensions IFL and IFR are controlled in the direction of resisting the roll. The load movement in the left and right direction becomes large, and the rear wheel suspensions IRL and IRR are controlled in a direction to escape from the roll, so the load movement in the left and right direction becomes small. When it is determined that the vehicle is turning left, the negative roll stiffness distribution target value - F gate is substituted for the stiffness distribution value FO, and the front wheel suspension IFL and IFR are similarly adjusted.
is controlled in a direction that resists roll, so the load shift in the left and right direction increases, and the rear wheel suspension IRL
, IRR is controlled in a direction away from the roll, so the load movement in the left and right direction becomes small.

In this way, step 2 as the turning state detection means M3
Based on the estimated lateral jerk 6RL as a turning transient state detected in the process of 90, a roll stiffness distribution target value FM is calculated in step 520, and the roll stiffness distribution control means M4 calculates the roll stiffness distribution target value FM by the processes of steps 530 to 710. Correction is performed to increase the roll stiffness distribution on the front wheel side in accordance with the roll stiffness distribution target value FM. Valve control processing is performed according to the corrected corrected control pressure amounts DEL, DFR, DRL, and DRR.

Next, in the valve control process shown in FIG. 9, the main gas chamber 4FL of each suspension IFL, IFR, IRL, IRR,
Gas is supplied and discharged to and from 4FR, 4RL, and 4RR.

That is, in step 810, the calculated torsion correction side i discharge pressure amounts DFL, DFR, DRL, DRR
In order to adjust the pressure of the main gas chambers 4FL, 4FR, 4RL, and 4RR based on the following formula, the high pressure reserve switching valve 26. 30, Leveling valve 42°44.46.
48 or discharge valve 5B, 60, 64.6
6 valve ◆On time tFL.

tFR, tRL, and tRR are calculated.

When high pressure reserve switching valves 26, 30 and leveling valves 42, 44, 46, 48 are on, that is, pressure increases, tFL = (aF/φ) - (DFL/PF) I) tF
R= (aF/φ') ・(DFR/PFH)tRL=
(aR/φ') ◆(DRL/PRII)tRR=(a
R/φ') ◆(DRR/PRH) discharge valve 5B, 60. 64. 66 ON, that is, when the pressure decreases, tFL = (bF/φ)・(D FL/ P FL) tF
R=(bF/φ)・(D FR/ P FR)tRL=
(bR/φ)・(D RL/P RL)tRR=(b
R/φ)・(D RR/ P RR) Here, aF/φ
, aR/φ is determined from the ratio PI/P2 between the tank pressure Pi (=PFH or PRH) on the high pressure side and the main gas chamber pressure P2 that receives gas from the high pressure tank, based on the map shown in Figure 20. . The high pressure side tank is the front wheel side or rear wheel side high pressure reserve tank 28.32, PFII is the pressure of the front wheel side high pressure reserve tank 2 days, and PRH is the pressure of the rear wheel side high pressure reserve tank 32. .

bF/φ and bR/φ are based on the map shown in Fig. 21,
It is determined from the ratio P2/P3 between the main gas chamber pressure P2 and the tank pressure P3 on the low pressure side from which gas is discharged from the main gas chamber. The low-pressure side tank is the low-pressure reserve tank 62, 6 on the front wheel side or rear wheel side.

Next, in the on-time correction calculation process of step 820, valve◆on-time tFL, tFR, tRL, tRR
Based on the following formula, the time during which the valve is actually driven (actual valve driving time) tFLU, tFR
U.

tRLLI, tRRU(tFLD, tFRD, t
RLD, tRRD) are calculated.

High pressure reserve switching valves 26, 30, leveling valve 42. 44. 46°, 48 on, that is, in the case of pressure increase tFLU=αF command tFL+βFL t FRU=αF◆tFR+βFR t RLU=αR◆tRL+βRL t RRU=αR◆tRR+βRR Distissage valve 5B, 60. 64. 66
On, that is, when the pressure decreases, t FLD=γF◆tFL+δFL t FRD=γF◆tFR+δFR t RLD=γR◆tRL+δRL t RRD=γR◆ tRR+δRR Here, αF, γF, αR1γR9βFL, βFR
, βRL.

βRR, δFL, δFR, δRL, and δRR represent predetermined coefficients.

Next, in step 830, the actual valve driving time tFL
tJ, tFRU, tRLU, tRRU(t
), tFLD, tFRD, t
Guard processing of RLD and tRRD (generally referred to as tD) is performed. This is to prevent the switching time of the valve from on to off or from off to on to become extremely short, and to protect the valve mechanism. That is, as shown in FIG. 22, when the actual valve driving times tU, tD with a duty of less than 30% are calculated, the actual valve driving times tu, to are set to zero, and the actual valve driving times tU, tD with a duty of less than 80% are calculated. When the driving times tu and to are calculated,
The actual valve drive times tu and to are fixed at a time corresponding to 80% duty.

Next, at the actual valve drive times tLl and tD guarded by step 840, valve 26. 30°42,
44t 46. 4EL 5B, 60. 64
．． The opening time of 66 is controlled.

In this way, when the air suspension control process is once completed and the process is restarted after a predetermined control period, step 10
A negative determination is made in step 3, and the process proceeds from step 110. The same process is repeated thereafter. In this embodiment, the control period is 100 m5, and the valve 26. 30.
42.44. 46°48.5B, 60, and 64.66 are duty-controlled within this 100m5. In other words, 1
Valve 26. 30. 4
2, 44. 46. 48. 5B, 60. 6
4. 66 is 100m until the next drive signal is issued.
The duty is controlled for 5 hours.

As described above, in this embodiment, each target pressure P FLM, P FRM, P RLM, P is repeatedly set at predetermined time intervals.
RRM is calculated, and each suspension IFL, IFR, IRL, IRR is calculated at predetermined time intervals according to the target pressure value.
Since the pressures of the main gas chambers 4PL, 4PR, 4RL, and 4RR are adjusted, smooth pressure control corresponding to actual roll changes is possible. Therefore, the driver's discomfort is eliminated and high steering stability can be achieved.

That is, in the case of relatively slow steering as shown in FIG. 23(A), the predicted lateral acceleration GRLM (double-dashed line) and the actual acceleration GRL are actually different as shown in FIG. 23(B).
(solid line). In this embodiment, as shown in (C), the target pressure is calculated in a predetermined cycle (for example, 100 m5) and a valve drive signal is output. Therefore,
As shown in (D), suspension IFL, IFR, I
RL, IRR gas chambers 4FL, 4FR, 4RL, 4R
Since the pressure R gradually increases stepwise in accordance with the degree of increase in the predicted acceleration GRLH, the roll angle can be kept extremely small as shown in (E). In this way, the vehicle becomes more stable and the handling stability is improved.

Further, as shown in FIG. 23(B), when the steered wheel 8 is operated and the estimated jerk 5RL is obtained, the roll stiffness distribution target value FM is calculated. Next, when it is determined that the steering is a right turn, the roll stiffness distribution target value FM is substituted into the stiffness distribution value FO, and each suspension IFL is
IFR, IRL.

The IRR is controlled in a direction that resists roll.

In a transient turning state, as shown in Fig. 24, the front axle side suspension IFL and IFR are designed to reduce the front wheel right side suspension force FFL compared to during steady turning, and to increase the front left side suspension force FFL compared to during steady turning. Since this is controlled, the load movement in the left and right direction on the front wheel side becomes larger than during steady turning. - side, rear wheel side suspension IR
L and IRR are controlled so that the rear wheel suspension force FRR is increased compared to that during steady turning, and the rear left suspension force FRL is decreased compared to that during steady turning, so that the load shift in the left and right direction on the rear wheel side is reduced. As a result, the steering characteristic of the vehicle becomes understeer. Furthermore, when it is determined that the turning is to the left, a negative roll stiffness distribution target value -FM is substituted for the stiffness distribution value FO, and similarly each suspension IF
L, IFR, IRL.

The IRR is controlled in a direction that resists roll.

Therefore, in a transient state of turning, the left-right load shift on the front wheels becomes larger than during steady turning, while the left-right load shift on the rear wheels becomes smaller than during steady turning, and the steering characteristics of the vehicle become understeer. . Furthermore, when the estimated jerk 6'RL becomes smaller, that is, when the turning steady state is reached,
The roll stiffness distribution target value FM becomes zero or a predetermined value, and the steering characteristics of the vehicle are switched to those in a steady state.

Therefore, as shown in Fig. 25, if the characteristics of the vehicle are such that the front wheel suspensions IFL and IFR are subject to roll, the dynamic characteristics of the yaw rate will be as shown by the broken line. By controlling the directional movement load, the front wheel suspensions IFL and IFR receive the roll, and as shown by the trajectory of the 0 mark, the responsiveness of the yaw rate increases and the responsiveness of the steering improves. In addition, when the turning is in a steady state, predetermined vehicle characteristics such as front and rear wheel suspensions IFL, IFR, IRL, and IRR shown by solid lines are applied.
It is possible to receive the roll and maintain the effectiveness of the rudder.

In addition, in the above embodiment, the air circuit AC and step 1
03 to 510. The processing from 810 to 840 corresponds to the processing by the attitude control means M2, the processing from step 290 corresponds to the processing from the turning state detection means M3, and the processing from steps 520 to 710 corresponds to the processing from the roll stiffness distribution control means. This corresponds to processing as M4.

In this embodiment, each gas spring 2PL,
l is controlled by supplying and discharging air to 2PR, 2RL, and 2RR, but the actuator is controlled by supplying and discharging hydraulic oil,
It is also possible to implement the method by controlling the posture.

Although the embodiments of the present invention have been described above, the present invention is not equally limited to these embodiments, and it goes without saying that the present invention can be implemented in various forms without departing from the gist of the present invention.

As described in detail above, the electronically controlled suspension system of the present invention has the effect of controlling the load shift in the left and right direction during a turning transient state and improving the steering response during turning. Qin.

[Brief explanation of the drawing]

Fig. 1 is a basic configuration diagram of the present invention, Fig. 2 is a schematic configuration diagram of an embodiment of an electronically controlled suspension device, Fig. 3 is an air circuit diagram of this embodiment, and Fig. 4 is an electrical circuit diagram of this embodiment. A block diagram showing the system configuration, FIG. 5 is a general flowchart of the control routine executed by the electronic control circuit of this embodiment,
Figure 6 is a flowchart of the feedforward calculation process 1, Figure 7 is a flowchart of the feed pump calculation process, Figure 8 is a flowchart of the corrected total pressure calculation process, and Figure 9 is a flowchart of the corrected total pressure calculation process. is a flowchart of valve control processing, FIG. 10 is a graph showing a map for calculating estimated lateral acceleration 6RL from steering angle θ and vehicle speed V, and FIG. The graph shown in Fig. 12 is the predicted acceleration G R
From LM to target pressure difference, 41PFLM, ΔPFRM, ΔP
A graph showing a map for determining RLM and ΔP RRM, shown in FIG. 13, is a feed forward gain of 1 and a feedback gain of 2 based on the difference between predicted acceleration GRLM and actual lateral acceleration GRL. A graph representing a map for finding 3 in ,
Figure 14 is a graph corresponding to the map for determining the target vehicle height based on vehicle speed V and mode, and Figure 15 is the actual longitudinal acceleration GF.
Graph corresponding to the map for determining the target pinch displacement XPM based on the actual lateral acceleration GRL, Figure 16 is the graph corresponding to the map for determining the target roll displacement XRM based on the actual lateral acceleration GRL, and Figure 17 is the actual longitudinal acceleration. A graph corresponding to the map for calculating the target Bi-Sochi displacement velocity pattern based on GFR,
Figure 18 is a graph corresponding to the map for determining the target roll displacement speed statement RM based on the distance dRL of the actual ship;
The figure is a graph corresponding to a map for calculating the roll stiffness distribution target value FM based on the estimated jerk dRL, and Figure 20 shows the relationship between the tank pressure P1 on the high-pressure side and the main gas chamber pressure P2 that receives gas from the high-pressure tank. The graph corresponding to the map that calculates the coefficients aF/φ and aR/φ based on the ratio Pi/P2, Figure 21 shows the main gas chamber pressure P2 and the pressure of the tank on the low pressure side that receives gas from the main gas chamber. Ratio P with P3
2/P3 is a graph corresponding to the map for calculating the coefficients bF/φ and bR/φ based on 3. Figure 22 is a graph corresponding to the map for calculating the output duty based on the actual valve drive times tu and to. Figure 23 is the timing chart of this embodiment, and Figure 24 is the suspension force FFL of this embodiment.
, FFR, FRL, and FRR over time. Figure 25 is a graph showing how the ratio between yaw rate YR and steering angle MA and the phase difference between the steering angle and yaw angle change depending on the steering speed. FIG. 26 is a graph showing the relationship between cornering power and inner and outer wheel loads. Ml...Suspension M2...Attitude control means M3...Turning state detection means M4...Roll rigidity distribution control means IFL, IFR, IRL, IRR...Suspension 2PL, 2PR, 2RL, 2RR...・Gas spring 26.
30... High pressure reserve switching valve 34, 36.
50. 52. 54゜56.70.72...Pressure sensor 42.44, 46.48...Leveling valve 5B,
60, 64.66...Discharge valve 80.8
2.84.86... Vehicle height sensor 90... Steering angle sensor 92... Acceleration sensor 93... Vehicle speed sensor 94... Door switch 95... Neutral switch 96... Throttle opening sensor 100...Electronic control circuit

Claims (1)

[Scope of Claims] An electronically controlled suspension having attitude control means for supplying and discharging fluid to suspensions provided corresponding to the wheels of the vehicle to control the attitude of the vehicle during turning with a predetermined front and rear wheel roll stiffness distribution. The device includes a turning state detection means for detecting a turning state from the running state of the vehicle, and a roll stiffness distribution on the front wheel side is increased in a turning transient state than in a steady turning state based on the turning state detected by the turning state detection means. An electronically controlled suspension device comprising: roll stiffness distribution control means for outputting a correction signal to the posture control means.