JPH0829653B2 - Active suspension - Google Patents

Description

TECHNICAL FIELD The present invention relates to an active suspension for a vehicle,
In particular, the present invention relates to improvement of an active suspension configured to generate an anti-roll moment against a roll of a vehicle body according to a lateral acceleration generated in the left-right direction of the vehicle.

[Conventional technology]

A conventional active suspension is disclosed in
Some are described in Japanese Patent Publication No. 61-182715.

This conventional example is a cylinder that is placed between each wheel and the vehicle body.
An actuator composed of a piston device or the like, a vehicle width direction acceleration detecting means for detecting acceleration in the vehicle width direction of the vehicle body, that is, a lateral direction, a roll angular acceleration detecting means for detecting roll angular acceleration of the vehicle body, and both accelerations An arithmetic and control unit is provided which controls each actuator based on the detection value of the detection means and increases or decreases the force acting between each wheel and the vehicle body via the actuator.

Then, the arithmetic and control unit determines, based on both input acceleration detection signals, that the directions of acceleration (centripetal acceleration) and roll angular acceleration in the vehicle width direction are relative to the center of gravity of an arbitrary portion above the center of gravity of the vehicle body. In view of the vehicle width direction acceleration detection value, it is determined that the vehicle is turning and if the determination means determines that the directions are different from each other. And the means for calculating the amount of load movement between the wheels and the determining means determine that they are in the same direction,
Assuming that the vehicle receives a crosswind, a means for calculating the load movement amount between the vehicle body and the corner wheels based on the roll angular acceleration detector, and a control for controlling each of the actuators according to the calculation result by each of the calculating means. And means.

[Problems to be Solved by the Invention]

However, in the conventional active suspension described above, the vehicle width is determined by determining whether the vehicle is turning or crosswind depending on the direction of the acceleration detection values of the vehicle width direction acceleration sensor and the roll angular acceleration sensor. Anti-roll control is performed based on the acceleration detection value of either the directional acceleration sensor or the roll angular acceleration sensor.For example, a leftward roll is received on the vehicle body due to the cross wind from the right when turning left. When it occurs, the acceleration detection value of the vehicle width direction acceleration sensor and the acceleration detection value of the roll angular acceleration sensor are in the same direction, and the actuator is controlled based on the acceleration detection value of the roll angular acceleration sensor. Actuator pressure decreases,
The pressure of the actuator on the turning inner wheel side will increase, rather than suppressing the roll due to inertial force, it will promote the roll over phenomenon to the contrary, and will also receive a cross wind from the left during left turn, which will affect the vehicle body. When a rightward roll occurs on the vehicle, the acceleration detection value of the vehicle width direction acceleration sensor and the acceleration detection value of the roll angular acceleration sensor are in different directions, and the actuator is controlled based on the acceleration detection value of the vehicle width direction acceleration sensor. As a result, there is no control corresponding to the amount of roll due to cross wind, and there is a problem that optimal anti-roll control cannot be performed.

In addition, if only the roll angular velocity is used to perform anti-roll control, a roll that can be actually detected after a cross wind is generated, and a control response delay occurs mainly due to a process delay. There is also a problem that control becomes extremely unstable, such as oscillation, depending on the relationship between gain and phase.

Further, if the roll flat control based on the vehicle width direction acceleration is performed almost completely, the roll angular acceleration due to the cross wind is a very small amount centered on zero, and there is a deviation in control in that direction. Since it can be positive or negative, there is also a problem that it is practically impossible to distinguish the inertial force from the cross wind by using the roll angular acceleration.

Therefore, the present invention has been made by paying attention to the problems of the above-mentioned conventional example, and is an active type capable of stabilizing the control system while exhibiting an accurate anti-roll effect against a crosswind during turning. Intended to provide suspension.

[Means for solving the problem]

In order to achieve the above object, an active suspension according to claim (1) is provided with an actuator inserted between a vehicle body and each wheel, and a lateral acceleration detecting means for detecting a lateral acceleration of the vehicle body. A roll displacement of a vehicle body in an active suspension including attitude change control means for controlling the actuators in a direction of suppressing the attitude change of the vehicle body based on a value obtained by multiplying a detected value of the lateral acceleration detecting means by a gain. The posture change control means includes a roll displacement detection means for detecting, and the posture change control means has a configuration for correcting the roll suppression control amount for each actuator based on the detection value of the lateral acceleration detection means by the roll displacement detected by the roll displacement detection means. It is characterized by that.

The active suspension according to claim (2) is
The attitude change control means is characterized by increasing the correction amount due to the roll displacement in accordance with the increase of the vehicle speed.

Further, the active suspension according to claim (3) is characterized in that the posture change control means increases the front and rear wheel distribution ratio of the correction amount due to the roll displacement in comparison with the front wheel side to the rear wheel side.

[Action]

In the active suspension according to claim (1), the attitude change control means for controlling the actuators interposed between the wheels and the vehicle body controls the roll displacement control means based on the detection value of the lateral acceleration detection means. Since the roll displacement detected by the method is used to correct the roll displacement, the roll displacement increases with respect to the cross wind, so the roll suppression control amount can be increased and corrected to deal with it. Can be eliminated.

Further, in the active suspension according to the second aspect, by increasing the correction amount due to the roll displacement in accordance with the increase in the vehicle speed, it is possible to ensure the steering stability during high speed running.

Further, in the active suspension according to claim (3), since the front / rear wheel distribution ratio of the correction amount due to the roll displacement is larger than that of the front wheel side to the rear wheel side, the steer characteristic of the vehicle is understeer. At the same time, the yaw moment force point due to the combined force of the cornering forces on the front wheel side and the rear wheel side is behind the force point of the lateral force due to the side wind, so that a yaw moment that changes the direction toward the leeward direction is generated. The cross wind stability is improved because it can be corrected by operation.

〔Example〕

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

1 to 10 are views showing an embodiment of the present invention.

In FIG. 1, 11FL, 11FR, 11RL and 11RR are active suspensions interposed between a vehicle body side member 12 and a wheel side member 14 that individually supports the wheels 13FL, 13FR, 13RL and 13RR, respectively. There are hydraulic cylinders 15FL to 15RR for attitude control and coil springs 16FL as actuators.
˜16RR and attitude control hydraulic cylinders 15FL to 15RR are provided with pressure control valves 17FL to 17RR for controlling the operating hydraulic pressure in response to only a command value from a control device 36 described later.

Here, in each of the attitude control hydraulic cylinders 15FL to 15RR, as shown in FIGS. 1 and 2, the cylinder tube 15a is attached to the wheel side member 14, and the piston rod 15c is provided with the piston 15b at the tip. Is attached to the vehicle body side member 12, and the hydraulic pressure of the hydraulic chamber 15d in the cylinder tube 15a is controlled by the pressure control valves 17FL to 17RR. In addition,
The piston 15b has a through hole 15d that communicates between the upper and lower hydraulic chambers 15d.
The cylinder tube 15a and the piston rod 15c are moved relative to each other due to the difference between the pressure receiving areas above and below the piston 15b.

Each of the coil springs 16FL to 16RR is mounted in such a manner that the piston rod 15c is wound between the body side member 12 and the cylinder tube 15a of the attitude control hydraulic cylinders 15FL to 15RR to support the static load of the vehicle body. are doing. In addition,
The coil springs 16FL to 16RR may have a low spring constant that only supports the static load of the vehicle body.

As shown in FIG. 2, each of the pressure control valves 17FL to 17RR has a cylindrical valve housing 21 and a proportional solenoid 22 integrally provided with the valve housing 21. A partition wall 21A having a valve seat 21c having a predetermined diameter is provided in the center of the valve housing 21.
The upper insertion hole 21U in FIG. 2 and the lower insertion hole 21L in FIG. 2 defined by the above are coaxially formed. Further, a fixed throttle 23 is provided above the insertion hole 21L and below the partition 21A by a predetermined distance, whereby a pilot chamber C is formed between the fixed throttle 23 and the partition 21A. A main spool 24 is disposed below the fixed throttle 23 in the insertion hole 21L so as to be slidable in the axial direction, and feedback chambers F _U and F _L are provided above and below the main spool 24, respectively. While being formed, the upper and lower ends of the main spool 24 are regulated by offset sprockets 25A and 25B arranged in the feedback chambers F _U and F _L , respectively. Then, the input port 21i, the control port 21n, and the drain port 21o are formed to communicate with the insertion hole 21L in this order,
The input port 21i is connected to the hydraulic pressure source 18 via the line pressure pipe 19.
Drain port 21o is connected to the discharge side of
Connected to the tank of the hydraulic supply source 18 via the control port 21n via the hydraulic pipe 27 to the hydraulic cylinders 15FL to 15R.
It is connected to the R pressure chamber 15d.

The main spool 24 includes a land 24a facing the input port 21i, a land 24b facing the drain port 21o,
These two ports 24a, includes a pressure chamber 24c comprising an annular groove formed between 24b, and a pilot passage 24d for communicating the feedback chamber F _L of the pressure chamber 24c and a lower side.

In the upper insertion hole 21U, a poppet 26 is axially slidably arranged with its valve portion facing the valve seat 21c. The poppet 26 allows the insertion hole 21U to move in two axial directions.
The chamber is defined, and the flow rate of the working oil flowing through the valve seat 21c, that is, the pressure in the pilot chamber C can be adjusted.

Further, the input port 21i is provided with an orifice 21 _S0
Is communicated with the pilot chamber C via a pilot passage 21s having the above, and the drain port 21o is communicated with the insertion hole 21U via a drain passage 21t.

On the other hand, the proportional solenoid 22 includes an axially slidable plunger 27, an actuator 27A fixed to the poppet 26 side of the plunger 27, and an exciting coil 28 for driving the plunger 27 in the axial direction. Has and this excitation coil
28 is appropriately excited by a command current I from the controller 36. As a result, the movement of the plunger 27 controls the position of the poppet 26 via the actuator 27A, and controls the flow rate passing through the communication hole 21A. When the pressures of the feedback chambers F _L and F _U are balanced while the pressure force of the proportional solenoid 22 is being applied to the poppet 26,
The spool 24 is in the neutral position, and the control port 21n is disconnected from the input port 21i and the drain port 21o.

Here, as for the relationship between the command current I and the control oil pressure P _C output from the control port 21n, as shown in FIG. 3, P _MIN is output when the command value I is near zero, and from this state When the command value I increases in the positive direction, the control hydraulic pressure P _C increases with a predetermined proportional gain K ₁ and is saturated with the line pressure P _L of the hydraulic pressure supply source 18.

The pressure control valves 17FL to 17RR are connected to the proportional solenoid 22.
Pressing force is applied to the poppet 26 by, and in a state where the pressure in the upper feedback chamber F _U and the lower feedback chamber F _L is balanced, the wheel, for example, upward sprung resonance frequency due to the convex portion passes the road surface When a vibration input of a relatively low frequency (or a downward vibration input due to passage through a recess) is transmitted, the hydraulic cylinder 15FL ~
The cylinder tube 15a of 15RR tries to move upward (or downward), and the pressure in the hydraulic chamber 15a increases (or decreases).

Thus, the pressure in the hydraulic chamber 15d is increased (or decreased), the hydraulic chamber 15a and the hydraulic pipe 27, a pressure control port 21n and the communication via the pilot passage 24d through the lower feedback chamber F _L accordingly Rises (or falls) and loses its balance with the pressure in the upper feedback chamber F _U , so that the spool 24 moves upward (or downward) and the input port 21i
And the control port 21n are closed (or opened) and the output port 21o and the control pressure port 21n are opened (or closed). Part of the pressure is discharged from the control pressure port 21n to the hydraulic pressure supply source 18 via the output port 21o and the hydraulic pressure pipe 20 (or from the hydraulic pressure source 18 to the input port 21i, the control pressure port 2).
The hydraulic pressure is supplied to the hydraulic chamber 15d via the 1n and the hydraulic pipe 27).

As a result, the pressure in the hydraulic chamber 15d of the hydraulic cylinders 15FL to 15RR is reduced (or increased), and the pressure increase in the hydraulic chamber 15d due to the upward vibration input (or the pressure decrease in the pressure chamber 15a due to the downward vibration input) is suppressed. Therefore, the vibration input transmitted to the vehicle body side member 14 can be reduced. At this time, since no restriction is provided in the drain pipe 20 between the output port 21o of the pressure control valves 17FL to 17RR and the hydraulic pressure supply source 18, a damping force should be generated when suppressing the upward vibration input. There is no.

In FIG. 1, 28H is a pressure control valve 17FL to 17RR.
And a high-pressure side accumulator for pulsation absorption connected in the middle of the hydraulic pipe 25 between the hydraulic source 24 and the hydraulic source 24, and 28L is a pressure control valve 17FL to 1FL.
A low pressure side accumulator that absorbs pressure fluctuations in the unsprung resonance frequency range that cannot be followed by the pressure control valves 17FL to 17RR that are connected to the hydraulic pipe 27 between the 7RR and the hydraulic cylinders 15FL to 15RR via the throttle valve 28V. is there.

Further, the vehicle body is provided with a lateral acceleration detector 35 for detecting lateral acceleration at a position slightly forward of the sprung center of gravity position. When the lateral acceleration of the vehicle is zero, the lateral acceleration detector 35 detects the lateral acceleration. Zero and a positive voltage proportional to the lateral acceleration when the lateral acceleration occurs during the right turn, and a negative voltage proportional to the lateral acceleration when the lateral acceleration occurs during the left turn are output as lateral acceleration detection values.

Further, between the vehicle body side member 12 and the wheel side member 14 that individually supports the front wheels 134FL, 13FR, hydraulic cylinders 15RL, 15RR are provided.
In parallel, relative displacement detectors 36L, 36R constituting a roll amount detecting means, such as a potentiometer, for detecting relative displacement between the two are arranged, and these relative displacement detectors 36L, 36R are arranged.
From the relative displacement detection value S _L , which is a voltage corresponding to the relative displacement amount,
S _R is output.

Furthermore, for example, a vehicle speed detector 37 is provided on the output side of a transmission (not shown), and the vehicle speed detector 37 outputs a vehicle speed detection value V having a voltage corresponding to the vehicle speed.

The lateral acceleration detected value of the lateral acceleration detector 35, a relative displacement detector 36L, relative displacement detected value S _L of 36R, S _R and the vehicle speed detection value V of the vehicle speed detector 37 is input to the control unit 38.

The control device 38, as shown in FIG.
The lateral acceleration detection value from 35 is supplied, and the front wheel side roll suppression gain adjuster 39F and the rear wheel side roll suppression gain adjuster are configured by an amplifier that amplifies the lateral acceleration detection value with predetermined gains (amplification degrees) Ky _F and Ky _R. and 39R, the relative displacement detected value of the relative displacement detectors 36L and 36R S _L and S _R is input, representative of the roll displacement based on the difference value obtained by subtracting the relative displacement detected value S _L from the relative displacement detected value S _R A roll angle calculating circuit 40 as a roll displacement detecting means for calculating the roll angle θ, a roll angle θ of the roll angle calculating circuit 40, and a vehicle speed detector.
The vehicle speed detection value V from 37 is input, and the gains Ks _F and Ks _R are changed as shown in FIG. 5 according to the vehicle speed detection value V. For example, a front wheel side roll correction gain adjuster 41F composed of a variable gain amplifier and With the rear wheel side roll correction gain adjuster 41R,
Attitude change control command by adding the front wheel roll suppression command value VR _F output from the front wheel roll suppression gain adjuster 39F and the front wheel roll rigidity command value VS _F output from the front wheel roll correction gain adjuster 41F Front wheel side adder 4 that outputs the value VC _F
2F and the rear wheel side roll restraining gain adjuster wheel side roll correction command value after outputted from the rear wheel side roll correction gain adjuster 41R and wheel side roll restraining instruction value VS _R after output from the 39R VS _R
Neutral voltage addition for adding the wheel side adder 42R after outputting the attitude change suppressing command value VC _R by adding a predetermined neutral voltage V _N to the attitude change suppressing command value VC _F from the front wheel side adder 42F bets Bowl 43
FL, a sign inverter 44F that multiplies the posture change suppression command value VC _R from the front wheel side adder 42F by -1, and a neutral voltage adder that adds a predetermined neutral voltage V _N to the output of the sign inverter 44F. 43FR, a neutral voltage adder 43RL for adding a predetermined neutral voltage V _N to the addition output of the rear wheel side adder 42R, and a sign inverter 44R for multiplying the addition output of the rear wheel side adder 42R by -1.
, A neutral voltage adder 43RR for adding a predetermined neutral voltage V _N to the output of the sign inverter 44R, and each neutral voltage adder 43FL to
The added output of 43RR is individually input, and the command currents I _{FL to} I _RR of the current proportional solenoids 22 of the pressure control valves 17 _{FL to} 17 _RR are controlled according to these, for example, a control valve composed of a floating type constant current circuit. It is provided with drive circuits 45FL to 45RR.

Here, the gains Ky _F and Ky of the front wheel side roll suppression gain adjuster 39F and the rear wheel side roll suppression gain adjuster 39R, respectively.
_As shown in FIG. 6, _R is selected so that Ky _F > Ky _R is always obtained in order to always obtain the understeer characteristic, and the front wheel side roll correction gain adjuster 41F and the rear wheel side roll correction gain adjuster 41R are selected. As shown in FIG. 5, each gain Ks _F and Ks _R of the vehicle speed detection value V has a predetermined set value V _S (for example, 50 km /
to zero when less than h), when the vehicle speed detecting value V is equal to or larger than a predetermined set value V _S is selected to increase while maintaining the relationship Ks _F> Ks _R in accordance with the increase in the vehicle speed detection value V ing.

Next, the operation of the above embodiment will be described. Now, it is assumed that the vehicle is traveling straight on a good road with no unevenness on the road and no cross wind. In this state, since the lateral acceleration does not act on the vehicle body, the value of the lateral acceleration detection value of the lateral acceleration position detector 35 becomes substantially zero, and no rolling occurs.Therefore, the relative displacement detector 36L arranged on the rear wheel side. , 36R relative displacement detection values S _L , S _R are equal. Therefore, the roll suppression command values VR _F and VR output from the front wheel side roll suppression gain adjuster 39F and the rear wheel side roll suppression gain adjuster 39R of the control device 38.
_{As R} also becomes substantially zero, the roll angle θ of the roll angle calculation circuit 40 becomes zero, and the roll correction command value VS output from the front wheel side roll correction gain adjustment value 41F and the rear wheel side roll correction gain adjuster 41R. _F and VS _R also has become a zero. Accordingly, the adder 42F and attitude change suppressing command value VC _F and VC _R also becomes zero is output from the 42R, the neutral voltage adder 43FL~43RR
Neutral voltage V _N is output as the command voltage V _FL ~V _RR from, since they are supplied to the control valve driving circuit 45FL～45RR, corresponding to the neutral voltage V _N to the proportional solenoid 22 of the pressure control valve 17FL~17RR command current I _N is supplied to hold, as the control pressure P _FL to P _RR of the pressure control valve 17FL~17RR is shown in FIG. 6, is held in a predetermined neutral pressure P _N, the vehicle body is zero rolling state To be done.

Therefore, in this state, as described above, due to the vibration input of a relatively low frequency input from the road surface through the wheels 13FL to 13RR, the pressure fluctuations in the pressure control chamber C of the pressure control valves 17FL to 17RR are caused. The vibration input of a relatively high frequency that is absorbed by the movement of the spool 24 and corresponds to the unsprung resonance frequency due to the fine unevenness of the road surface is absorbed by the throttle valve 28V and the accumulator 28L, and the vibration transmissibility to the vehicle body is increased. It is possible to reduce the weight and secure a good riding comfort.

When the steering wheel is turned to the right by turning the steering wheel from the straight running state, as shown in FIG. 7, the vehicle body tends to roll to the right with a roll angle θ when viewed from the front side. At this time, as shown in FIG. 8, one wheel of the vehicle will be described. Here, the mass of the vehicle is M, and the effective area of the hydraulic cylinder 15FL is A.

Considering the linear range of the characteristic of the pressure control valve 17 shown in FIG. 3, P = K ₁ · I (7)

On the other hand, from FIG. 8, M ₂ = P · A + K (x ₁ −x ₂ ) ... (8), and the command current I becomes I = Ky ... (9).

Therefore, by substituting the equation (7) into the equation (8) and rearranging, M ₂ + K (x ₂ −x ₁ ) = K ₁ · IA ··· (10).

Here, when the unsprung displacement x ₁ is zero (x ₁ = 0) and is expressed by a transfer function in the form of the response of the sprung displacement x ₂ to the command current I, Therefore, if expressed as a transfer function in the form of response to lateral acceleration, Becomes

On the other hand, in vehicles that do not perform anti-roll control,
As shown in the figure, the roll motion of the vehicle body when lateral acceleration is applied, the roll moment of inertia is J, the roll angle is θ,
When the distance between the center of gravity and the roll center is H, the spring constant is K, and the tread is L, it can be expressed by the following equation.

In equation (13), when expressed by a transfer function in the form of response to lateral acceleration, Further, since x ₂ = Lθ / 2, this can be calculated from the above (13).
Substituting it in the equation and expressing it as a transfer function in the form of response to lateral acceleration, Becomes

Therefore, the expression (12) represents the response of the hydraulic system of the present invention, and the expression (15) represents the roll motion with respect to the lateral acceleration. When the two are compared, the denominators are both quadratic and equivalent.

Therefore, it can be understood that the roll motion can be dynamically suppressed by the hydraulic system of the present invention by appropriately setting the gain Ky of the expression (12).

Therefore, when the vehicle is turning to the right, the lateral acceleration detection value of the lateral acceleration detector 35 is a positive value, which is respectively the front wheel side roll suppression gain adjuster 39F and the rear wheel side roll suppression gain adjuster 39R. Is supplied to.

The gain regulator 39F and the gain lateral acceleration detected value Ky _F and Ky _R multiplied by front and rear wheels roll reduction command value VR _F and VR _R front wheel side adder 42F and the rear-wheel-side adder output from the 39R Since it is directly supplied to the neutral voltage adders 43FL and 43RL as the front wheel side attitude control command value VC _F and the rear wheel side attitude control command value VC _R via 42, the front wheel and rear wheel attitude control commands are supplied from these adders 43FL and 43RR. since the values VC _F and VC _R to the neutral pressure setting voltage V attitude control command _N obtained by adding the voltage V _FL and V _RL is supplied to the control valve driving circuits 45FL and 45RL, the pressure control valve by these driving circuits 45FL and 45RL 17FL And 17RL proportional solenoids 22 are supplied with command currents I _FL and I _{RL that} are higher than the neutral current I _N , which causes pressure control valves 17 _FL and 17 _FL.
The control pressures P _FL and P _{RL of RL} become higher than the neutral pressure P _N , as shown in FIG.

On the other hand, the pressure control valves 17FR and 17RR on the right side are provided with the front wheel and rear wheel roll suppression command values VR _F and VR _{R, respectively.}
Since they are supplied to the neutral voltage adders 43FR and 43RR via 4R, their control pressures P _FR and P _RR are lower than the neutral pressure P _N as shown in FIG.

Therefore, the pressure in the hydraulic chamber 15d of the left hydraulic cylinders 15FL and 15RL increases. Therefore, in the pressure receiving area of the upper hydraulic chamber of the piston 15b and the pressure receiving area of the lower hydraulic chamber, the pressure receiving area of the upper hydraulic chamber is smaller than the pressure receiving area of the lower hydraulic chamber by the cross-section integral of the piston rod. Therefore, the thrust that is obtained by multiplying the area difference between the two by pressure acts upward, and this thrust causes the hydraulic cylinder 15FL on the left side to move.
And 15RL can generate a cylinder biasing force that resists the contracting force contracted by the rolls, and can exert the anti-roll effect that maintains the vehicle body in the zero roll state.

Also, the hydraulic chamber 15d of the right hydraulic cylinders 15FR and 15RR
The pressure is reduced, which reduces the thrust for urging the piston 15b upward for the same reason as above, and controls the urging force so as not to promote the extension force of the roll.

In this way, the hydraulic cylinder
Results 15FL~15RR is controlled so as to exert anti-roll effect, the body maintains a zero roll state, thus maintaining the relative displacement detectors 36L, relative displacement detected value S _L of 36R, S _R also substantially equal Therefore, the roll angle θ output from the roll angle calculation circuit 40 also keeps zero.

When the vehicle receives a cross wind from the right during this right turn, this cross wind causes a roll in the same direction as the roll due to the centrifugal force at the time of right turn, which is leftward when viewed from the rear side. . When this cross wind roll is generated,
Since the relative displacement detection value S _R of the relative displacement detector 36R becomes larger than the relative displacement detection value S _L of the relative displacement detector 36L, the difference value between the detection values S _R and S _L from the roll angle calculation circuit 40. A positive roll angle θ based on is output. At this time, when the vehicle speed detection value V of the vehicle speed detector 37 is less than the predetermined set value V _S is the roll stiffness gain adjuster 41F, the gain of 41R K
Since s _F and Ks _R are zero, these gain adjusters 41F and
Roll correction command value VS _F output from the 41R, VS _R maintains a zero roll angle when the vehicle speed detecting value V is equal to or larger than a predetermined set value V _S, then the respective gain adjuster 41F, the 41R θ
The roll correction command values VS _F , VS _{R, which} are obtained by multiplying the gains Ks _F , Ks _R by, are output to the front wheel side and rear wheel side adders 42F and 42R, and the roll suppression command value VR _F based on the positive lateral acceleration detection value. , Added to VR _R. As a result, the adder 42F, attitude control command value is outputted from the 42R VC _F, VC _R roll correction command value VS _F, is larger by the correction VS _R component, generates a normal stabilizer equivalent roll reaction force Since the roll rigidity is increased, the roll displacement of the vehicle body can be suppressed to be small, and the course can be prevented from being disturbed by the roll steer, and the ground camber can be reduced to secure the tire grip force, the wheel and the vehicle body. It is possible to suppress changes in suspension geometry due to stroke changes between the two and to stabilize the driver's field of view, which makes it possible to move the vehicle sharply and stably when making turns or changing diagonal lines. Focusing on the yaw moment due to the side wind and the cornering force generated by the tire in the cross wind condition at the time, the front wheel side roll correction gain adjuster 41F In Ks _F and gain of rear wheel side roll correction gain adjuster 41R
Since Ks _R is selected as Ks _F > Ks _R , as shown in FIG. 10, the cornering force CF _F on the front wheel side becomes larger than the cornering force CF _R on the rear wheel side (CF _F
> CF _R ), the resultant force point NSP of both cornering forces is on the rear side of the vehicle with respect to the force point AC of the lateral force due to the side wind.
The vehicle will be subjected to a yaw moment such that the front wheels will turn leeward. Such a yaw moment is
It is an input in a stable direction for the vehicle, can be easily corrected by steering operation, and can ensure crosswind stability. Incidentally, the gain Ks _R of the front-wheel-side roll correction gain adjuster 41F gain Ks _F and rear wheel side roll correction gain adjuster 41R, when selected as Ks _F <Ks _R, force application point NSP of the resultant force of the cornering force crosswind The lateral force of the vehicle is on the front wheel side of the point AC, and the rear wheel side of the vehicle receives the yaw moment that flows downwind.To correct this yaw moment, the steering must be operated in reverse. Stability decreases.

Further, when the vehicle speed detection value V is equal to or higher than the predetermined set value V _S and when a crosswind is received from the left side while traveling in the right turn, the crosswind rolls the vehicle body to the right and rolls in the direction opposite to the roll in the right turn. However, at this time, the relative displacement detector
The relative displacement detection value S _L of 36L becomes larger than the relative displacement detection value S _R of the relative displacement detector 36R, and therefore the roll angle θ of the roll angle calculation circuit 40 becomes a negative value. Therefore,
With the adders 42F, 42R, the roll suppression command values VR _F , VR _{R are} to be corrected by subtracting the roll correction command values VS _F , VS _R.
Attitude control command value VC _F, VC _R roll stiffness command value VS _F, VS _R min becomes smaller, the left pressure control valves 17FL and 17RL in accordance with this
Together with the control pressure P _FL and P _FL of drops, the stabilizer function by increasing the control pressure P _FR and P _RR of the right pressure control valves 17FR and 17RR, generates a roll reaction force hydraulic cylinder 15FL~15RR It is possible to suppress the roll due to the side wind from the left.

Similarly, when the vehicle turns left, the lateral acceleration detection value of the lateral acceleration detector 35 becomes a negative value, and accordingly, as shown in FIG. 6, the control pressures of the left pressure control valves 17FL and 17RL are controlled. P _FL , P _RL is lower than the neutral pressure P _N , and the control pressures P _FR , P _RR of the pressure control valves 17FR, 17RR on the right side are higher than the neutral pressure P _N , so that the anti-roll effect and the stabilizer function are simultaneously exerted. be able to.

Further, even when a roll due to a centrifugal force is transiently generated in the vehicle body during turning, a difference in the relative displacement detection values S _L and S _R is generated. Therefore, in order to suppress this, a roll counter equivalent to a normal stabilizer is used. Since a force is generated to increase the roll rigidity, the roll displacement of the vehicle body can be suppressed to be small.

Further, when the vehicle body rolls due to a lateral wind while the vehicle is traveling straight at a vehicle speed equal to or higher than the predetermined set value V _S , the lateral acceleration detection value of the lateral acceleration detector 35 is substantially zero.
The stabilizer function can be exerted by correcting the roll suppression command value based on the difference value between the relative displacement detection values S _L and S _R of the relative displacement detectors 36L and 36R, so that the roll due to cross wind can be suppressed. .

In the above embodiment, the case of obtaining a large selection to understeer characteristic gain Ky _F of the front wheel side roll restraining gain adjuster 39F the gain Ky _F of the gain Ky _R of the rear wheel side roll restraining gain controller 39R However, the present invention is not limited to this, and both gain adjusters 39L and 39R
Is composed of a variable gain amplifier, and these gains are varied while keeping the total gain constant based on the steering angle detection value of the steering angle detector.
_By setting _F ≤ Ky _R , it is possible to improve the turning performance of the vehicle as a neutral steer characteristic or oversteer characteristic, and then as the steering angle increases, Ky _F > Ky _R , and it is possible to secure running stability as an understeer characteristic. , The gains Ky _F and Ky _R of both gain adjusters 39F and 39R are equalized to have a neutral steer characteristic, and the understeer characteristic at the time of turning is obtained by, for example, steering the rear wheel steering mechanism in phase with the front wheels. Good.

Further, in the above embodiment, the case where the lateral acceleration generated in the vehicle is detected by the lateral acceleration detector 35 has been described, but the present invention is not limited to this, and the vehicle speed V and the steering angle δ are detected as the vehicle speed detector and the steering angle detection, respectively. A lateral acceleration detection device (see JP-A-62-293167, rearward) for detecting a true lateral acceleration generated in the vehicle by performing predetermined arithmetic processing based on these may be provided. In this case, the control can be performed based on the true lateral acceleration that is not affected by the roll of the vehicle body, so that the control accuracy can be improved and the control system is substantially an open loop system. Therefore, there is an advantage that there is no possibility of causing self-excited vibration.

Further, in the above-described embodiment, the case where the lateral acceleration detector 35 outputs a positive acceleration detection value when turning to the right and a negative acceleration detection value when turning to the left has been described, but the present invention is not limited to this. From 35, a positive predetermined value is output when the lateral acceleration is zero, a positive value higher than the predetermined value when turning to the right, and a positive value lower than the predetermined value when turning to the left, respectively. Even when a subtracter for subtracting the neutral set voltage V _N is applied instead of the neutral voltage adders 43FL to 43RR of 38, the same effect as that of the above embodiment can be obtained.

Furthermore, in the above embodiment, the case where the relative displacement detectors 36L and 36R disposed on the front wheel side are applied as the roll displacement detection means has been described, but the present invention is not limited to this, and the relative displacement detectors are disposed on the rear wheel side. You may.

Further, in the above embodiment, the case where the hydraulic cylinder is applied as the actuator has been described.
Needless to say, the present invention is not limited to this, and other fluid pressure cylinders such as a pneumatic cylinder can be applied.

Further, in the above embodiment, the case where the gains of the roll stiffness gain adjusters 41F and 41R are changed according to the vehicle speed has been described, but the present invention is not limited to this, and a constant gain may be applied. .

Furthermore, in the above embodiment, the case where the roll suppression gains Ky _F and Ky _R are set to constant values has been described, but the present invention is not limited to this, and the steering angle can be changed to the steering angle without changing these total gains. It may be changed based on, in this case, as the neutral steer characteristic during straight running, the oversteer characteristic at the start of turning, and the understeer characteristic during turning,
The turning ability at the start of turning can be improved.

〔The invention's effect〕

As described above, according to the active suspension according to claim (1), the attitude change control means for controlling the actuator interposed between the vehicle body and each wheel is provided with the roll displacement based on the lateral acceleration. Since it is configured to correct by, the anti-roll effect can be exerted with high accuracy and the vehicle body can be maintained in the zero roll state without being affected by the side wind disturbance during turning of the vehicle, and the ground camber can be reduced. It is possible to secure tire grip power, suppress suspension geometry change due to stroke change between wheels and vehicle body, and stabilize the driver's field of view. It is a snappy and stable vehicle when turning, changing diagonal lines, etc. The movement of the roll is possible and the roll rigidity is controlled by detecting the roll displacement instead of the roll angular acceleration, so it is stable without oscillation. Effect that can be controlled can be obtained.

Further, according to the active suspension of the second aspect, the correction amount due to the roll displacement is increased according to the increase of the vehicle speed, so that the crosswind stability of the vehicle at the time of high speed traveling can be further improved.

Further, according to the active suspension according to claim (3), since the front and rear wheel distribution ratio of the correction amount by the roll displacement is made larger than that of the front wheel side to the rear wheel side, cornering when a side wind is received. The force applied point of the resultant force can be set to the rear side of the vehicle with respect to the side force applied by the lateral wind, and the change in the traveling direction of the vehicle due to the lateral wind can be corrected by the normal steering operation, thus the steering stability is improved. Can be improved.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a schematic configuration diagram showing an embodiment of the present invention, FIG. 2 is a sectional view showing an example of a pressure control valve applicable to the present invention, and FIG. Is a characteristic diagram showing the relationship between the command current of the pressure control valve and the output pressure, FIG. 4 is a block diagram showing an example of the control device applicable to the present invention, and FIG. 5 is the relationship between vehicle speed and roll rigidity gain. FIG. 6 is a characteristic diagram showing the relationship between the lateral acceleration and the control pressure of the pressure control valve, and FIGS. 7 to 10 are explanatory diagrams for explaining the operation of the present invention. In the figure, 11FL to 11RR are active suspensions, 12 is a vehicle body side member, 13FL to 13RR are wheels, 14 is a wheel side member, and 15FL to 15FL.
RR is a hydraulic cylinder (actuator), 17FL to 17RR are pressure control valves, 35 is a lateral acceleration detector, 36L and 36R are relative displacement detectors, 37 is a vehicle speed detector, 38 is a control device, and 39F is a front wheel roll suppression gain. Regulator, 39R is a rear wheel side roll suppression gain adjuster, 40 is a roll angle calculation circuit (roll displacement detection means), 41F is a front wheel side roll rigidity gain adjuster, 41R is a rear wheel side roll rigidity gain adjuster, 42F is Front wheel side adder, 42R rear wheel side adder, 43FL to 43RR neutral voltage adder, 44F, 44R
Is a sign inverter, and 45FL to 45RR are control valve drive circuits.

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Yosuke Akatsu 2 Takara-cho, Kanagawa-ku, Yokohama, Kanagawa Nissan Motor Co., Ltd. (72) Masaharu Sato 2 Takara-cho, Kanagawa-ku, Yokohama, Kanagawa Nissan Motor Co., Ltd. ( 72) Inventor Kensuke Fukuyama 2 Takaramachi, Kanagawa-ku, Yokohama, Kanagawa Nissan Motor Co., Ltd. (56) Reference JP 62-34808 (JP, A) JP 62-152910 (JP, A) JP Sho 60-25811 (JP, A)

Claims (3)

[Claims] 1. An actuator inserted between a vehicle body and each wheel, lateral acceleration detecting means for detecting lateral acceleration of the vehicle body, and a value obtained by multiplying a detected value of the lateral acceleration detecting means by a gain. Based on the above, in an active suspension including attitude change control means for controlling each of the actuators in a direction of suppressing the attitude change of the vehicle body, a roll displacement detection means for detecting a roll displacement of the vehicle body is provided, and the attitude change control means is An active suspension characterized in that a roll restraint control amount for each of the actuators based on a detection value of the lateral acceleration detecting means is corrected by a roll displacement detected by the roll displacement detecting means. 2. The active suspension according to claim 1, wherein the attitude change control means increases the correction amount due to the roll displacement in accordance with an increase in vehicle speed. 3. The posture change control means is characterized in that the front and rear wheel distribution ratio of the correction amount due to the roll displacement is made larger than that of the front wheel side to the rear wheel side. Active suspension.