JPH0719850Y2 - Active suspension - Google Patents

Description

[Detailed Description of the Invention] [Industrial field of application] The present invention relates to an active suspension, and in particular, to a fluid pressure cylinder interposed between a sprung and an unsprung mass, and an operating pressure of the fluid pressure cylinder. And a pressure control valve for controlling the command value according to a changeable command value, and controlling the command value according to the posture change of the vehicle body.

[Conventional technology]

As an active suspension of a vehicle, for example, there is one described in Japanese Patent Application No. 61-137875 previously proposed by the present applicant.

This prior application detects lateral acceleration or longitudinal acceleration generated in the vehicle body, and controls the output pressure of the pressure control valve, that is, the working pressure of the hydraulic cylinder between the vehicle body and the vehicle body, by the command value according to the detected value. , Roll rigidity or pitch rigidity is controlled.

Among these, the pressure control valve is composed of a pilot type proportional electromagnetic pressure reducing valve. That is, the pressure control valve includes a three-way spool valve inserted between the hydraulic power source and the hydraulic cylinder, and a proportional solenoid that controls the spool position of the three-way spool valve according to a command value. . Of this,
The hydraulic oil from the hydraulic pressure source is supplied to the supply port (input port in the previous application) of the three-way spool valve, the return port (output port in the previous application) is connected to the drain side of the hydraulic source, and the output port (previous application) The input / output port) is connected to the hydraulic cylinder. This pressure control valve feeds back the control pressure output from the output port to the pressure control chamber of the spool, and balances the pressure based on this feedback pressure with the thrust of the proportional solenoid, and is supplied from the hydraulic power source. The line pressure is reduced according to a predetermined command value and is supplied to the hydraulic cylinder as control pressure.

On the other hand, this pressure control valve has a large flow rate of hydraulic oil from the hydraulic cylinder side when the vehicle passes through a relatively gentle uneven part and the vibration input in the sprung resonance frequency range (about 1 Hz) is in the hydraulic cylinder. Is returned (or supplied) to the hydraulic pressure source side,
As a result, the output port and the return port (or the supply port) of the spool valve communicate with each other, and the working oil is returned (or supplied) to the hydraulic pressure source to absorb the pressure fluctuation.

[Problems to be solved by the device]

However, in the above-mentioned pressure control valve for the active suspension, since the spool movement amount is usually manufactured with a sufficient margin to provide versatility, a large flow rate in the sprung resonance frequency range from the road surface side. When there is a changeable oscillating input, the stroke amount of the spool becomes a large value that is almost free according to the change in the flow rate, so the throttling effect is especially effective for the oscillating input in the sprung resonance frequency range. It is small and does not generate sufficient damping force.
For this reason, there is an unsolved problem that the change in the control pressure at the time of vibration in such a situation is small, and therefore the effect of suppressing the vibration transmitted from the road surface side to the vehicle body side cannot be sufficiently exerted.

The present invention has been made in view of such an unsolved problem, and in particular, with respect to a vibration input of a relatively low frequency from the road surface side, without affecting the vibration frequency vs. gain characteristic, It is an object of the present invention to provide an active suspension equipped with a pressure control valve capable of generating a larger damping force.

[Means for Solving the Problems]

In order to achieve the above object, the present invention comprises a fluid pressure cylinder interposed between a sprung portion and an unsprung portion, and a pressure control valve for controlling the working pressure of the fluid pressure cylinder. A pressure adjustment for reciprocating the spool according to a differential pressure between a spool slidable through an insertion hole in the valve housing and a pilot pressure which is variable based on a change in the attitude of the vehicle body and a pressure output by the pressure control valve. In the active suspension including a mechanism, the pressure control valve limits the stroke range of the spool to a predetermined range necessary for roll control or pitch control of the vehicle body by making the stroke range of the spool shorter than its full stroke range. Means are provided.

[Action]

In this invention, the stroke amount of the spool of the pressure control valve in the roll control and the pitch control of the posture change suppression control is usually larger than that in the bounce control, and more generally, the road surface side, that is, the fluid pressure cylinder. The focus is on the fact that the stroke amount during the pressure adjusting operation by the vibration input in the sprung resonance frequency range from the side is larger than that in the roll control or the pitch control, and the flow rate change is also large.

Therefore, when such vibration is applied from the road surface side, in the present invention, the stroke regulating means regulates the stroke amount of the spool of the pressure control valve to be smaller than the full stroke range thereof, which results in a large throttle. An effect is obtained, and a large damping force is generated by it. For this reason, the pressure output by the pressure control valve changes greatly, and the vibration transmitted to the vehicle body side is significantly reduced. Further, posture change suppression control including roll control can be performed without any trouble.

〔Example〕

(First Embodiment) An embodiment of the present invention will be described below with reference to FIGS. 1 to 10. This embodiment shows a hydraulic active suspension that controls roll of a vehicle.

First, the overall configuration will be described with reference to FIG. 2 (the figure shows only one of the hydraulic systems for the four wheels of the vehicle).

In the figure, 10 is an active suspension, 12 is a wheel,
Reference numeral 14 is a wheel side member, and 16 is a vehicle body side member.

The active suspension 10 includes a hydraulic pressure source 18 that supplies a hydraulic pressure of a predetermined line pressure, an accumulator 20 for accumulating pressure installed downstream of the hydraulic pressure source 18, and a pressure control valve installed downstream of the accumulator 20. 22, a hydraulic cylinder 24 as a fluid pressure cylinder interposed between the vehicle body side member 16 and the wheel side member 14, and a vehicle body static cylinder disposed between the vehicle body side member 16 and the cylinder tube 24a of the hydraulic cylinder 24. A coil spring 30 that supports a load and a posture change suppression control device 32 that controls the pressure control valve 22 are provided. The posture change suppression control device 32 controls the lateral acceleration sensor 34 for detecting the lateral acceleration of the vehicle body, the longitudinal acceleration sensor 35 for detecting the longitudinal acceleration, and the pressure control valve 22 based on the lateral acceleration or longitudinal acceleration detection signal. And a controller 36 for controlling. Further, a pressure chamber L of the hydraulic cylinder 24, which will be described later, is provided.
Is communicated with the accumulator 40 via the throttle valve 38 in order to absorb the vibration in the unsprung resonance frequency range from the road surface side.

The hydraulic power source 18 uses a vehicle engine as a rotational drive source and has a hydraulic pump that pressurizes and discharges hydraulic oil in a tank, and outputs hydraulic oil of a predetermined line pressure based on the discharge pressure of the hydraulic pump. To do.

As shown in FIG. 1, the pressure control valve 22 is composed of a pilot-operated proportional electromagnetic pressure reducing valve. The pressure control valve 22 includes a valve mechanism section 22A including a three-way spool valve and a proportional mechanism for controlling the valve mechanism section 22A. And a solenoid 22B.

Of these, the valve mechanism portion 22A has a valve housing 42, and a substantially cylindrical insertion hole 44 is formed in the center portion of the valve housing 42. At a predetermined position of the insertion hole 44, a partition wall 46 having a communication hole 46A having a predetermined diameter and used for variable pilot pressure is provided laterally in the axial direction, whereby the insertion hole 44 is divided into two parts. Among these, the poppet 48 axially biased by the proportional solenoid 22B is inserted into the insertion hole 44 on the proportional solenoid 22B side.
Are slidably arranged. On the other hand, in the insertion hole 44 on the opposite side, a fixed throttle 50 is provided laterally at a predetermined distance from the partition wall 46, and the pilot chamber is provided between the fixed throttle 50 and the partition wall 46.
PR is formed. The fixed throttle 50 suppresses the disturbance of the pressure P _p in the pilot chamber PR due to the displacement of the main spool 52, which will be described later. A plurality of holes may be provided around the periphery.

Further, the main spool 52 described above is disposed in the insertion hole 44 on the fixed throttle 50 side so as to be slidable in the axial direction thereof, and the valve housing 42 has a supply port communicating with the insertion hole 44.
54s, an output port 54c, and a return port 54r are formed respectively. The supply port 54s is connected to the hydraulic source 1 via the hydraulic piping.
8 is connected to the hydraulic oil supply side, the return port 54r is connected to the drain side of the hydraulic pressure source 18 via the hydraulic pipe, and the output port 54c is connected to the pressure chamber L of the hydraulic cylinder 24 described later via the hydraulic pipe. Has been done.

A pilot-side pressure chamber _FU and an output-side pressure chamber _FL are formed at axial positions of the insertion hole 44 facing the main spool 52. Then, offset pressures 60U and 60L are arranged in both pressure chambers F _U and F _L , respectively.
In the pressure parallel state of both pressure chambers F _U and F _L , the sliding position of the main spool 52 is centered (a pressure chamber 52c described later is closed as shown), and the offset amount becomes zero.

The main spool 52 includes a land 52a facing the supply port 54s, a land 52b facing the return port 54r, and an annular groove-shaped pressure chamber 52c formed between the lands 52a and 52b.
When, it is formed by a feedback passage 52d communicating with the pressure chamber 52c and the output-side pressure chamber F _L. The feedback passage 52d is provided with an orifice Pc for a damper as shown in the figure.

On the other hand, the pilot chamber PR communicates with the pilot-side pressure chamber F _U via the fixed throttle 50, and the pilot supply passage
It communicates with the supply port 54s via PP. In addition, the insertion hole 4
The poppet 48 side of 4 communicates with the return port 54r through the pilot return passage PT. Furthermore, the both ends of the poppet 48 in the insertion hole 44 and the pilot return passage PT are mutually
It is connected via the communication passage PR.

Then, the the pilot supply passage PP multistage orifice Qp is provided for regulating the pilot flow, multi-stage orifice Pr is provided to prevent the disturbance of the pilot pressure P _p is the pilot return path PT.

Therefore, the opening area of the communication hole 46A is changed by adjusting the sliding position of the poppet 48, and the pilot supply passage P
The flow rate of the hydraulic oil flowing through P, the pilot chamber PR, the communication hole 46A, and the pilot return passage PT, that is, the pilot pressure P _p of the pilot chamber PR is adjusted, whereby the pilot side pressure chamber F _U
Pressure is controlled. When the pressures in the pressure chambers F _U and F _L are different, the pressure difference is applied to the main spool 52.
Moves reciprocally in the axial direction, and the pressure adjusting operation for inflowing or discharging the hydraulic oil in the pressure chamber 52c is transiently performed. When the pressure in the pressure chamber 52c (that is, the control pressure P _c of the output port 54c) and the pilot pressure P _p are balanced, the spool position shown in FIG. 1 is always maintained in the subsequent stationary state when the hydraulic cylinder 24 is fixed. Is taken and the control pressure P _c set equal to the pilot pressure P _p is held.

On the other hand, the proportional solenoid 22B includes a cylindrical housing 62 formed integrally with the valve housing 42 and a cylindrical housing 62.
Insertion hole 64 formed in the center of 62 and communicating with the aforementioned insertion hole 44
And a plunger 66 slidably provided in the axial direction of the insertion hole 64, and an exciting coil 68 for driving the plunger 66 in the axial direction. Plunger 66 actuator
The tip of 66A is in contact with the poppet 48. Also 66B
Is an insertion hole. In addition, the plunger 66 is inserted into the communication hole 66B.
An orifice may be provided for damping the above.

Then, the exciting coil 68 is supplied with an exciting current I _s as a command value from a posture change suppression control device 32 described later. Therefore, the exciting coil 68 is excited according to the value of the exciting current I _s , and the actuator 66A of the plunger 66 biases the poppet 48 downward in FIG. 1 according to the propulsive force thereof.

On the other hand, on the inner wall of the insertion hole 44 on the side of the main spool 52, a stopper serving as a stroke restricting means is provided at a position apart from both end positions of the centered main spool 52 by a predetermined distance.
70U and 70L are fixed respectively. Therefore, the main spool 52 can freely move according to the pressure imbalance when the stroke amount is equal to or less than the distance d between the stoppers 70U and 70L, but when the distance exceeds the distance d.
The end of the main spool 52 contacts the stoppers 70U, 70L,
Further displacement is impossible.

Next, a method of calculating the above-mentioned separation distance d in this embodiment will be described below.

First, the differential pressure between the upstream side and the downstream side of the pressure control valve 22,
For example, when flowing from the supply port 54s to the output port 54c, the differential pressure between the supply pressure and the output pressure, when flowing from the output port 54c to the return port 54r, the differential pressure between the output pressure and the return pressure, passing through the pressure control valve 22. The flow rate to be performed and the spool displacement at that time will be described.

Assuming that the differential pressure between the upstream and downstream of the spool is ΔP, the spool passage amount is Q, and the spool opening area is A, then Q = CA [(2 / ρ) ΔP] ^1/2 (1) C: flow coefficient (= 0.65) ρ: Density of hydraulic oil (= 8.8 to 8.9 × 10 ^-7 kgfs ² / cm ⁴ ).

The relationship between the spool displacement X and the opening area A is A≈αX (2) α: Spool opening area constant (= 6.25 mm ² / mm). Then, the relational expression of Q≈CαX [(2 / ρ) ΔP] ^1/2 (3) is obtained from the equations (1) and (2).

From the equation (3), the relationship between Q, ΔP and X in the case of the present embodiment is shown in FIG.

Next, the flow rate required for roll control and pitch control will be estimated.

First, a linearized model of the load system from the pressure control valve outlet to the hydraulic cylinder is obtained.

1. Basic formula The basic formula is as follows, considering the inertia of the oil column and ignoring the compressibility of oil.

M _{L v} / A _L = (P _v -P _a -R _L Q _v ) A _L …… (4) Q _v ＝ Q _s ＋ Q _b …… (5) P _a −P _b ＝ R _b Q _b …… (6) Q _a ＝ A _a V _a …… (7) P _b ＝ K _b ∫ Q _b dt …… (8) Where, M _L : oil column mass, A _L : pipe cross-sectional area, R _L : pipe Flow resistance, R _b : Damping valve flow resistance, K _b : Accumulator constant, A _a : Actuator (hydraulic cylinder) pressure receiving area,
V _a : piston speed, P: pressure, Q: flow rate (see Fig. 5).

2. Transfer Matrix When the above basic equation is Laplace transformed, the relationship between the valve outlet pressure P _v and the flow rate Q _v and the actuator pressure P _a and the piston speed V _a can be expressed using the transfer matrix as follows.

From the above, the relationship between the pressure and the flow rate at the pressure control valve outlet and the cylinder is obtained.

Further, in order to simplify the calculation, assuming that the characteristic of the pressure control valve is the first-order lag of the response characteristic of the control pressure with respect to the command pressure, P _sig = (1 + Ts) P _v ...... (11) P _sig : Command pressure P _v : Pressure control valve outlet pressure.

From equations (9), (10), and (11), assuming the roll control or pitch control, the cylinder will not move in a fixed state (in the equation)
When V _a = 0), the relation between the control pressure P _a in the cylinder and the valve passage flow rate Q _v with respect to the command pressure P _sig is obtained, and the following equation is obtained.

FIG. 6 shows the response (Bode diagram) to the command pressure P _sig obtained from the equation (12).

In addition, the flow rate through the pressure control valve at the time of the same response as equation (12)
When calculated by the equation (3), it becomes as shown in Fig. 7 (pressure amplitude 5
0 ± 20kgf / cm ² o'clock).

FIG. 6 shows the responsiveness in roll control and pitch control required in this embodiment, and FIG. 7 shows the flow rate required at that time.

According to this, the maximum flow rate required for roll control and pitch control is about 6 l / min.

Next, we will estimate the flow rate through the pressure control valve when the cylinder is vibrated.

First, to simplify the calculation, assume that the pressure control valve is a fixed throttle (for example, suppose that the Pco orifice represents the vibration characteristic of the pressure control valve). Or (P _v ′, Q _v ′ indicates the pressure and flow rate on the upstream side of the pressure control valve).

From equations (14), (15), (9) and (10), P _v ′ is constant (the pressure at the upstream Pc port immediately upstream of the Pco orifice in the pressure control valve is ideally adjusted to a constant pressure). , The following equation is obtained.

However, in the calculation, V _a is positive on the pressure side.

The equation (16) expresses the cylinder vibration characteristics required in this embodiment. R'which becomes equivalent to the required performance at a cylinder vibration speed of 0.3 m / s is selected and calculated to obtain the eighth characteristic. It is calculated as shown in the figure.

When the flow rate through the pressure control valve at the vibration characteristic shown in FIG. 8 is calculated by the equation (17) (at 0.3 m / s vibration), it becomes as shown in FIG.
According to this, the maximum flow rate during vibration is about 8 l / min.

Now, when arranging here, * The flow rate passing through the pressure control valve is about 6l / min at the time of roll and pitch control, and about 8l / min at the time of cylinder vibration (at 0.3m / s) * Upstream of the pressure control valve. differential pressure side and a downstream side, when the maximum flow rate occurs, roll, means for pitch control at about 70 kgf / cm ² cylinder pressure Futoki about 63.5kgf / cm ² above "about 70 kgf / cm ^2" is roll and pitch since the control time control pressure is controlled to about 50 ± 20kgf / cm ^2, next ^{2 30~70kgf} / cm as control pressure boosts from 30 kgf / cm ^2, the flow from the supply pressure side to the control pressure side, when the time becomes the maximum flow rate Since it can be considered, 100−30 = 70 kgf / cm ² . Conversely, 70kgf /
voltage reduction from cm ^2, flows into the pressure side returns from the control pressure side, be considered that the time becomes the maximum flow rate, 70-0 = 70kgf / cm ²
Therefore, the differential pressure when the maximum flow rate is generated is considered to be about 70 kgf / cm ² .

Meaning of "about 63.5kgf / cm ^2", since 0.3 m / s pressurized Futoki 1Hz damping constant of about 220kgfs / m, the pressure fluctuation at the time (attenuation constant ÷ receiving area × piston speed) is 220 ÷ It becomes 4.9 × 0.3 ≒ 13.5kgf / cm ² . Furthermore, this performance is neutral pressure from the pressure fluctuations when about 50 kgf / cm ² is controlled, the control pressure 50-13.5kgf / cm ² in the stretching step, shrinkage in the process control pressure 50 + 13.
It will be 5 kgf / cm ² .

Then, in the elongation step, the pressure flows from the supply pressure side to the control pressure side, so that eventually 100− (50−13.5) = 63.5 kgf / cm ² . On the contrary, in the shrinking process, the pressure flows from the control pressure side to the return pressure side, so it becomes 50 + 13.5-0 = 63.5kgf / cm ² and the differential pressure when the maximum flow rate is generated is 63.5kgf / cm ²
Becomes

FIG. 10 shows the spool operating point at the above-mentioned flow rate and differential pressure. The spool displacement during roll and pitch control is approximately 0.4 m
It becomes m (point A in the figure). Also, the spool displacement during vibration when obtaining the damping constant (220 kgfs / m) calculated above is
It can be seen that it becomes about 0.5 mm (point B in the figure).

After all, for roll control and pitch control, it can be said that the spool displacement should be about 0.4 mm. And, it is necessary to have a damping constant of 220 kgfs / m to 300 kgfs as a characteristic at the time of vibration.
It is said that the larger the 1 m / m attenuation is, the better the 1 Hz attenuation is. For example, if the spool operation is restricted within the spool stroke range required for roll control and pitch control, the damping constant can be made considerably large.

That is, if the distance d in FIG. 1 is determined so that it can be stroked ± 0.4 mm from the neutral position of the spool, damping at low frequency vibration can be achieved without impairing roll control and pitch control performance. Become.

The specifications used in this calculation are shown below.

C: Flow coefficient 0.65 ρ: Density 8.8 × 10 ^-7 [kgfs ² / cm ⁴ ] α: Spool opening area constant 6.25 [mm ² / mm] T: Time constant 50 × 10 ^-3 M _L : Oil column mass 1.02 × 10 ^-4 [kgfs ² / cm] A _L : Pipe cross-sectional area 0.58 [cm ² ] R _L : Pipe resistance 1.31 × 10 ^-3 [kgfs / cm ⁵ ] R _b : Damping valve resistance 0.407 [kgfs / cm ⁵ ] K _b : Accumulator constant 2.59 [kgf / cm ⁵ ] R ': Control valve equivalent resistance 0.1 [kgfs / cm ⁵ ] A _a : Cylinder rod cross-sectional area 4.91 [cm ² ] Here, in the pressure control valve 22, the proportional solenoid is used. twenty two
B, the poppet 48, bulkhead 46, the fixed throttle 50, the pilot supply passage PP, the pilot return path PT, the pilot chamber PR, the pressure chamber F _U, F _L, 52c, which configures the main part of the greater pressure regulating mechanism in the feedback passage 52d There is.

Returning to FIG. 2, the hydraulic cylinder 24 has the cylinder tube 24a, and the piston 24c biased by the piston rod 24b is slidable in the axial direction in the cylinder tube 24a. A pressure chamber L separated by a piston 24c is formed below the cylinder tube 24a.

The controller 36 is configured to include a microcomputer, calculates lateral acceleration and longitudinal acceleration based on the detection signals of the lateral acceleration sensor 34 and the longitudinal acceleration sensor 35, and sets the lateral acceleration and longitudinal acceleration to a predetermined value. The value of the exciting current I _S as a command value is set within a predetermined range from I _{SNIM to} I _SMAX (see FIG. 3) by performing calculations such as multiplication with a gain constant, and the exciting current I _S is proportional to the proportional solenoid. It is designed to supply to 22B.

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

When the ignition switch is turned on, the posture change suppression control device 32 operates, and the hydraulic pressure source 18 is driven by the rotation of the engine as described above to supply a predetermined line pressure.

The posture change suppression control device 32 detects the lateral acceleration and the longitudinal acceleration acting on the vehicle body as described above, and performs a predetermined calculation for setting the command value. As a result, the excitation current I _S is set to a value near its neutral value I _SN during steady-state running where the vehicle is traveling straight on a good road at a constant speed. I _S is set to a value higher (for example, the outer ring side) and a lower value (for example, the inner ring side) than the neutral value I _SN .

Now, assuming that the vehicle is running normally as described above and there is no road surface input to the hydraulic cylinder 24, the exciting current
Assume that I _S is set to its neutral value I _SN . This energizes the proportional solenoid 22B, causing the plunger 6
6 biases the poppet 48 with a thrust of a predetermined value, and the flow resistance by the pocket 48 and the communication hole 46A is set to its neutral value. In other words, the pilot pressure P _p is also set to its neutral value, the pressure of the pilot pressure chamber F _U and the output-side pressure chamber F _L are compared.

At this time, if the pressures of the pressure chambers F _U and F _L are in equilibrium (that is, the control pressure P _c , that is, the pressure of the pressure chamber L of the hydraulic cylinder 24 and the pilot pressure P _p are equal), the main spool
The position of 52 continues the closed state of FIG. 1, and the control pressure P _c is also held at the neutral pressure P _CN .

For example, the pressure output side pressure chamber of the pilot pressure chamber F _U F _L
If the control pressure P _c is higher than the pilot pressure P _p , the main spool 52 moves in the direction of the pilot side pressure chamber F _U (upper side in FIG. 1) and the output port 54c
And the return port 54r communicates. Therefore, hydraulic fluid is returned to the port 54r side return via the pressure chamber 52c, the pressure of the output-side pressure chamber F _L decreases. Thereby, too low its pressure becomes lower than the pressure of the pilot pressure chamber F _U, go to the main spool 52 is now opposite direction (lower side in FIG. 1), between the supply port 54s and the output port 54c To communicate. Thereafter, after a transient state in which this cycle is repeated, the pressures in both pressure chambers F _U and F _L are balanced, the main spool 52 takes the closed position in FIG. 1 again, and the control pressure P _c becomes the neutral pressure P _CN . Set within a very short time.

When the pressure in the pilot side pressure chamber F _U is higher than the pressure in the output side pressure chamber F _L , that is, when the control pressure P _c is lower than the pilot pressure P _{p, the} pressure changes in the opposite way and the same pressure adjustment is performed. Be seen.

Further, assuming that the exciting current I _S is increased from the neutral state where the vehicle height value is in the proper range, the pilot pressure P _p is set high accordingly. Therefore, the pressure equilibrium between the two pressure chambers F _U and F _L is lost again, and the control pressure P _c is increased in accordance with the pilot pressure P _p , that is, the value of the exciting current I _S , through the same transient state as described above. It On the other hand, if the exciting current I _S is reduced, the control pressure P _c is reduced accordingly.

Therefore, in this embodiment, when the change of the control pressure P _c with respect to the change of the exciting current I _S is shown, the static characteristic of FIG. 3 is obtained. Here, P _CMAX is the maximum control pressure substantially equivalent to the line pressure of the hydraulic power source 18, and P _CMIN is the minimum control pressure.

In this way, in each pressure control valve 22, the hydraulic cylinder 24 corresponding to the control pressure P _c according to the value of the commanded excitation current I _S
Is supplied to. Therefore, the pressure chamber L of each hydraulic cylinder 24
Then, for example, a biasing force that opposes the roll is individually generated, whereby the roll rigidity is controlled, and a change in the posture of the vehicle body is appropriately suppressed.

At this time, the separation distance d between the stoppers 70U and 70L in the pressure control valve 22 is set to the maximum value of the stroke amount that can be taken by the normal roll control and the pitch control as described above, so that the main spool 52 is excited. It can operate as commanded by the current I _S.

On the other hand, it is now assumed that the pressures in both pressure chambers F _U and F _L are in equilibrium and in a steady state. In this state, the vehicle passes through a comparatively gentle uneven road, and the pressure chamber L of the hydraulic cylinder 24 starts from the road surface side.
It is assumed that there is a vibration input in the sprung resonance frequency range (about 1 Hz). This vibration input becomes a pressure fluctuation corresponding to it and is transmitted to the pressure chamber 52c of the pressure control valve 22. The flow rate change due to this pressure fluctuation is usually larger than that during pressure regulation by the exciting current I _S.

Due to this vibration input, the pressure in the output side pressure chamber F _L temporarily rises (or falls) and becomes higher (or lower) than the pressure in the pilot side pressure chamber F _U. Is biased to greatly move the pressure chamber F _U side so that the main spool 52 described above (or F _L side), the output port 54c and return port (or supply port 54s and the output port 54c) is communicated with The hydraulic oil is temporarily returned (or supplied) to the hydraulic pressure source 18. At this time, even if the main spool 52 tries to stroke beyond the set range d, it cannot be done by the stoppers 70U and 70L, so a damping force is generated due to the throttle effect due to the stroke restriction, and the control pressure P _c changes greatly, Fluctuations in the bounce direction of the vehicle are largely absorbed. In this way, the exciting current
Regardless of the value of I _S , it exerts a significant damping effect on the vibration input around 1 Hz, and greatly suppresses the fluctuation around 1 Hz in the bounce direction of the vehicle caused by the vibration transmitted from the road surface side to the vehicle body side. As a result, deterioration of riding comfort is accurately prevented.

Further, when the above-mentioned vibration input from the road surface side is of a high frequency corresponding to the unsprung resonance frequency due to the fine unevenness of the road surface, the pressure fluctuation of the pressure chamber L of each hydraulic cylinder 24 causes the throttle 38 to move. It is transmitted to the accumulator 40 via. As a result, similarly, the vibration transmissibility to the vehicle body can be significantly reduced.

(Second Embodiment) Next, a second embodiment of the present invention will be described with reference to FIGS. Here, the same reference numerals are used for the same components as in the first embodiment.

In the second embodiment, a spring mounted for offset is used as the stroke restricting means.

This will be described in detail with reference to FIG. 11. Instead of the stoppers 70U, 70L in the first embodiment, the pilot side pressure chamber
Springs 80U and 80L are separately installed in the F _U and output side pressure chambers F _L. The stroke-spring reaction force characteristics of the springs 80U and 80L are as shown in FIG. That is, up to the displacement amount D ₁ corresponding to the maximum stroke of the main spool 52 required for normal roll control, the straight line has a slope α ₁ proportional to the displacement amount D, and when the displacement amount D ₁ is exceeded, the slope sharply increases. It has a non-linear characteristic that increases with large α ₂ (> α ₁ ).

The spring characteristics of FIG. 12 described above are set as shown below.

According to the first embodiment, the spool stroke should be restricted to about 0.4 mm. That is, in FIG. 12, D ₁ is
It should be 0.4 mm. In addition, the spring of the spool has an equivalent damping effect, but the spool diameter was 13 mm experimentally.
In the case of, α ₁ has been confirmed to be 10 kgf / mm ² (if the spool diameter is 10 mm, α ₁ is 5 kgf / mm ² ).

Further, as to α ₂ , as understood from the first embodiment, α ₂ = ∞ may be satisfied.

As this method, there is a method in which the spring is in close contact at D ₁ = 0.4 mm.

Further, as shown in the calculation data of the first embodiment, if it is desired to set the damping constant during vibration at low frequency to a certain value between 220 and 300 kgfs / m, for example, α ₁ <α ₂ < If set to a value of ∞, it can be set arbitrarily.

As this method, for example, there is a method using a so-called non-linear spring having the characteristic shown by the one-dot chain line in FIG.

As described above, the springs 80U and 80L in this embodiment are
In a region with a small spring constant, it functions as an offset setting, and in a region with a large spring constant, it almost functions as a stopper.

Other configurations and operations are the same as those of the first embodiment.

Therefore, when the main spool 52 is displaced so as to exceed the stroke amount in the normal roll control or pitch control, the displacement is almost restricted by the springs 80F and 80L, and the throttle effect is obtained. The same thing as the above-mentioned first embodiment can be obtained. Further, in this embodiment, since the springs 80U and 80L also serve as the stroke restricting means, it is possible to save the labor of providing the stopper as compared with the first embodiment.

In addition, although each of the above-described embodiments has described the case of roll control and pitch control, this may be control that combines bounce control, for example.

The fluid pressure cylinder in this invention may be a pneumatic cylinder.

[Effect of device]

As described above, in the present invention, the pressure control valve is the active suspension in which the stroke restricting means for restricting the stroke range of the spool to the predetermined region necessary for the roll control and the pitch control of the vehicle body is provided. That is, there is a vibration input of a relatively low frequency corresponding to the sprung resonance frequency range from the fluid pressure cylinder side, and a working fluid with a larger flow rate than that in roll control or pitch control flows into the pressure control valve or from the pressure control valve. Even if it flows out, the stroke amount of the spool during pressure regulation does not exceed the maximum stroke value during roll control, so a large damping force can be generated by the throttling effect due to stroke regulation, and the pressure during vibration is reduced. A large change in the pressure output by the control valve can be taken. Practical control that can significantly reduce vibration in the sprung resonance frequency range that is transmitted from the road surface side to the vehicle body side without affecting the frequency characteristics of the pitch control). Active suspension can be provided.

[Brief description of drawings]

FIG. 1 is a schematic sectional view showing a first embodiment of the present invention, and FIG.
FIG. 1 is a block diagram showing the outline of an active suspension for a vehicle according to the first embodiment, and FIG. 3 is an exciting current I _S of a pressure control valve versus a control pressure.
Graphs showing static characteristics of P _c , FIGS.
FIG. 11 is a graph for explaining an example of calculating the stopper position in the embodiment, FIG. 11 is a schematic sectional view showing the second embodiment of the present invention, and FIG. 12 is a graph showing the spring reaction force characteristics of the spring in the second embodiment. Is. In the figure, 18 is a hydraulic power source, 22 is a pressure control valve, 22B is a proportional solenoid, 24 is a hydraulic cylinder, 42 is a valve housing, 44 is an insertion hole, 46 is a partition wall, 48 is a poppet, 50 is a fixed throttle, and 52 is a main. Spool, 52c pressure chamber, 52d feedback passage,
70U and 70L are stoppers, 80U and 80L are springs that also serve as stroke control means, PP is a pilot supply passage, PT is a pilot return passage, PR is a pilot chamber, F _U is a pilot side pressure chamber, and F _L is an output side pressure chamber. is there.

Claims (3)

[Scope of utility model registration request] 1. A fluid pressure cylinder interposed between a sprung member and an unsprung member, and a pressure control valve for controlling an operating pressure of the fluid pressure cylinder. The pressure control valve is inserted into a valve housing. A spool slidable in the hole, and a pressure adjusting mechanism for reciprocating the spool according to a pressure difference between a pilot pressure that is changed based on a change in the attitude of the vehicle body and a pressure output from the pressure control valve are provided. In the active suspension, the stroke range of the spool is set to the pressure control valve.
An active suspension characterized by being provided with a stroke restricting means which is shorter than the full stroke range and is restricted to a predetermined region required for roll control and pitch control of the vehicle body. 2. The active suspension according to claim 1, wherein the stroke restricting means comprises stoppers provided at predetermined end positions of the insertion hole to stop sliding of the spool. 3. The stroke restricting means is a spring disposed at both ends of the spool, and the spring is provided when the amount of movement of the spool exceeds a predetermined area required for roll control or pitch control. The active suspension according to claim 1, wherein the constant is set to be extremely high.