JPH11316611A - Pressure compensating valve - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a pressure compensating valve capable of arbitrarily changing pressure compensation characteristics. SOLUTION: The pressure applied to a first pressure receiving part 21 energizes so as to expand the area of an opening between an inlet port 24 and an outlet port 25, the pressure applied to a second pressure receiving part 22 and a third pressure receiving part 23 energizes so as to decrease the said opening area. This valve is provided with a main valve 20 for respectively applying pressure P0 of pressurized oil to flow into the inlet port 24 to the first pressure receiving part 21 and applying pressure P3 of a load 5 driven by the pressurized oil flowing out of the outlet port 25 to the second pressure receiving part 22, and a control pressure generating means 7B for applying control pressure P1 reducing the pressure P0 at the inlet port 24 to the third pressure receiving part 23.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

[0001] The present invention relates to a pressure compensating valve used in a hydraulic system.

[0002]

2. Description of the Related Art FIG. 16 shows a hydraulic drive device disclosed in Japanese Patent Application Laid-Open No. 1-247805. The variable displacement type hydraulic pump A of this hydraulic drive device includes a low pressure side hydraulic cylinder D via a pressure compensating valve B and a directional control valve (operating valve) C.
It is connected to the. This pump A is provided with a pressure compensating valve B '.
And the high-pressure side hydraulic cylinder D 'via the direction control valve C'.
It is connected to the. The hydraulic pump A is provided with an actuator E for changing the displacement volume and a flow regulating valve F for controlling the actuator E.

The shuttle valve G detects the load pressure on the higher side of the load pressures generated by driving the cylinders D and D 'as the maximum load pressure _PLS , and uses the maximum load pressure _PLS as the flow control valve F As pilot pressure.

The flow control valve F is connected to the discharge pressure P _{P of the} pump A.
Is always higher than the maximum load pressure P _LS .

In this hydraulic drive system, the cylinders D and D 'are operated together by the simultaneous operation of the direction control valves C and C'. At this time, the pressure compensating valve B adjusts the amount of oil supplied to the cylinder D so that the difference between the inlet pressure and the outlet pressure of the directional control valve C is kept constant. The amount of oil supplied to the cylinder D 'is adjusted so that the difference between the inlet pressure and the outlet pressure becomes constant.

According to the above-described hydraulic drive system having the pressure compensating valves B and B ', there is a disadvantage that the pressure oil is concentrated and supplied to the light load side cylinder among the cylinders D and D' of the operating valve. Is prevented.

According to Japanese Patent Laid-Open No. 1-247805, when the amount of operation of the directional control valves C and C 'is large, a large amount of pressure oil is supplied to the low-pressure side hydraulic cylinder D and the discharge pressure of the pump A is increased. Decrease. In this case, the pressure compensating valve B
Does not reach the compensation differential pressure, the pressure compensation valve B does not perform the pressure compensation operation. That is, the pressure compensating valve B remains open.

When the pressure compensating valve B does not perform the pressure compensating operation, the amount of the pressure oil supplied to the low-pressure hydraulic cylinder D is not limited, so that the pressure oil is not supplied to the high-pressure hydraulic cylinder D '. This high-pressure side hydraulic cylinder D 'is not driven. Therefore, the operator must operate the direction control valve C in the small opening direction to limit the flow rate to the low-pressure side hydraulic cylinder D.

[0009] differential pressure P _P in this situation to prevent from occurring, the pressure P _P and the maximum load pressure P _LS of the pressure oil above the hydraulic drive system is discharged from the hydraulic pump A
A differential pressure detecting device H for detecting a -P _LS, differential pressure P _P -P _LS and the control force setting device I for setting a control force f _C based on the relationship shown in FIG. 17, the control force setting device I And an electromagnetic valve J that operates in response to the output signal.

The control force f _C is given by the following equation.

F _C = f−α (P _P− P _LS ) where f: springs b, b ′ of the pressure compensating valves B, B ′.
Pressing force α: Constant

When the differential pressure P _P -P _LS becomes equal to or less than the predetermined differential pressure P _m shown in FIG. 17, the solenoid valve J applies pressure oil corresponding to the control force f _C to the pressure receiving portions of the pressure compensating valves B and B ′. Act.

[0013] Thus the pressure compensating valve B, the spring b is the spool of the B ', the control force f _C against the pushing force f of b'added. This force f _C increases the flow resistance of the pressure compensating valves B and B ′ and increases the discharge pressure of the pump A, so that the pressure oil is also supplied to the high-pressure hydraulic cylinder D ′.

The above-mentioned cylinders D and D 'serve as working machines of construction machines (for example, booms, arms and buckets of hydraulic shovels).
In some cases, it may be desirable to change the pressure compensation characteristics of the pressure compensating valves B and B 'in order to improve workability depending on the working mode of the working machine.

In view of the above, the above-mentioned publication discloses a technique in which the throttle amount is made different for each pressure compensating valve, and the pressure difference between the front and rear of the directional control valves C and C 'can be made different as appropriate. That is, in this technique, the solenoid valve J is connected to the pressure compensating valves B and B '.
And the control force f _C for the pressure compensating valves B and B ′ is individually adjusted by the solenoid valves J. Therefore, the throttle amounts of the pressure compensating valves B and B 'are individually changed, that is, the differential pressures before and after the directional control valves C and C' are made different from each other.

A state in which the required flow rate is distributed irrespective of the load is particularly called a fully compensated state.

[0017]

However, the above-mentioned conventional apparatus has the following problems. When the mechanism that generates the control force f _C for changing the pressure compensation characteristics breaks down, pressure compensation may not be performed. Also, the differential pressure detecting device 2
After the differential pressure is detected by 1H, the solenoid valve J is driven through a calculation, so that the response is poor.

In view of such circumstances, an object of the present invention is to provide a pressure compensating valve which can change the pressure compensating characteristics arbitrarily, and has a high responsiveness and high reliability.

[0019]

According to a first aspect of the present invention, an opening area between an inlet port (24) and an outlet port (25) is increased by a pressure acting on a first pressure receiving portion (21). And the second pressure receiving portion (22)
The pressure acting on the pressure port (24) and the pressure acting on the third pressure receiving part (23) operate to reduce the opening area, and the pressure (P ₀ ) of the pressure oil flowing into the inlet port (24) is reduced by the first pressure. The pressure (P ₃ ) of the load (5) driven by the pressure oil flowing out from the outlet port (25) acts on the second pressure receiving portion (22). A main valve (20) and the third pressure receiving portion (2
Control pressure generating means for applying a pressure (P ₀₎ the control pressure the pressure was reduced to (P ₁₎ of the inlet port 3) (24) (7
B).

According to the first aspect, the control pressure (P
_Since the pressure compensation characteristic changes according to the magnitude of ₁ ), a desired pressure compensation characteristic can be obtained by changing the control pressure (P ₁ ).

The control pressure (P ₁ ) obtained by reducing the pressure of the inlet port (24) is applied to the third pressure receiving portion (2) of the main valve (20).
Since to act on 3), by an amount the pressure of the inlet port (24) corresponding to the variation amount when the variation control pressure (P ₁₎ also fluctuates. Therefore, the pressure compensation characteristic is not affected by the pressure fluctuation of the inlet port (24) of the main valve (20).

According to a second aspect, in the first aspect, the control pressure generating means (7B) includes a fixed throttle (36) and a variable throttle valve (30) connected in series with each other. It is characterized in that pressure oil from the pump side is made to flow through the variable throttle valve (30), and the oil pressure generated at the connection point between them is taken out as the control pressure (P ₁ ).

According to the second aspect, the variable throttle valve (3
The desired control pressure (P ₁ ) can be set by changing the throttle amount of 0).

Therefore, for example, when the load (5) is a hydraulic cylinder, if the pressure compensating valve according to the second aspect of the present invention is provided in the oil path on the bottom side and the oil path on the head side of the hydraulic cylinder, respectively, By individually setting the pressure compensation characteristics of the pressure compensation valves, it is possible to make the operation when the hydraulic cylinder expands and the operation when the hydraulic cylinder contracts differently.

According to a third aspect of the present invention, in the second aspect, the variable throttle valve (30) reduces the opening area between the inlet port (32) and the outlet port (33) by the elastic force of the spring (24). It operates so as to increase the opening area between the inlet port (32) and the outlet port (33) by the pilot pressure (P ₂ ) supplied from the pilot pressure supply means (7C).

According to the third aspect of the present invention, the pressure compensating function can be maintained even when the pilot pressure supply means (7C) fails and no longer generates pilot pressure. That is, when the pilot pressure disappears, the inlet port (32) and the outlet port (3) of the variable throttle valve (30)
Since 3) between is closed, so that the third pressure in the inlet port to the pressure receiving portion (23) (24) of said main valve (20) acts as a control pressure (P _1). Therefore, the pressure compensation function cannot be changed, but the pressure compensation function is maintained.

The fourth invention has a main valve (20) and a variable throttle valve (30) integrally connected to the main valve (20). The main valve (20) has an inlet port (24). A spool (S) provided with a valve part (66) for shutting off communication between the valve part (66) and a pushing part (67) for applying a pressing force to the valve part (66); A first pressure receiving portion (21) formed and receiving a hydraulic pressure for operating the spool (S) in the communication direction; and a first pressure receiving portion (21) formed on the spool (S) for operating the spool (S) in the shut-off direction. A second pressure receiving portion (22) and a third pressure receiving portion (23) for receiving the hydraulic pressure, and an inlet port (24) of the main valve (20) and an inlet port (3) of the variable throttle valve (30).
2) The first oil passage (37, 113, 11)
8, 119) and the first oil passages (37, 113, 11).
8, 119) with a fixed aperture (36, 113a,
119a) and the outlet port (2) of the main valve (20).
5) and a second oil passage (40) for communicating the outlet port (33) of the variable throttle valve (30) with the second oil passage (40). ) From the outlet port (33) to the outlet port (2) of the main valve (20).
Check valve (39,
120) and the inlet port (3) of the variable throttle valve (30).
2) and a third oil passage (38) for communicating the third pressure receiving portion (23) with the second pressure receiving portion (2).
It is characterized in that a load pressure (P ₃ ) is applied to 2).

According to the fourth aspect, the variable throttle valve (30) is integrally connected to the main valve (20), and the first oil passages (37, 113, 118, 119) and the second oil Road (4
0) and the third oil passage (38) are provided inside, so that the structure can be made compact.

According to a fifth aspect, in the fourth aspect, the load pressure (P _L ) of the load (5) driven by the pressure oil flowing out from the outlet port (25) of the main valve (20) is compared with another load pressure (P _L ). Load pressure comparison means (29) for comparing the load pressure of the load and applying the larger load pressure (P ₃ ) to the second pressure receiving portion (22) is incorporated in the main valve (20). It is characterized by.

According to the fifth aspect, since the load pressure comparing means (29) is incorporated in the main valve (20), a more compact structure can be realized.

According to a sixth aspect, in the fourth or fifth aspect, an electromagnetic proportional pressure control valve (5) for applying a variable pilot pressure to a spool (88) of the variable throttle valve (30).
0), the electromagnetic proportional pressure control valve (50) is provided with a spool (94) of the variable throttle valve (30).
8), the operating force of the spool (88) is applied to the spool (9) via a spring (54).
4) It is characterized in that it is arranged to provide mechanical feedback.

According to the sixth aspect, the pressure control valve (5
When the spool (88) of the variable throttle valve (30) is operated based on the pilot pressure supplied from (0), the operating force of the spool (88) is fed back to the spool (94) of the pressure control valve (50). Is done. Therefore, the dynamic characteristic of the spool (88) of the variable throttle valve (30) is improved, and more accurate pressure compensation becomes possible.

According to a seventh aspect of the present invention, in any one of the fourth to sixth aspects, the valve portion (6) of the spool (S) is provided.
6) and the push portion (67) are integrated.

According to the seventh aspect, the spool (S)
And the number of parts is reduced.

In an eighth aspect based on the seventh aspect, the first oil passage (113, 118, 1) is provided in the spool (S).
19), and the first oil passages (113, 118, 1) are formed.
19) is provided with the fixed stop (113a, 119a).

According to the eighth aspect, the first oil passage (1)
13, 118, 119) is provided on the spool (S) of the main valve (20), so that the first oil passage (1
13, 118, 119). Therefore, processing of the main valve (20) becomes easy.

In a ninth aspect based on the eighth aspect, the check valve (12) is provided in the first oil passage (118, 119).
0) is provided.

According to the ninth aspect, there is no need to provide the check valve (120) in the body. Therefore, processing of the body is further facilitated.

[0039]

FIG. 1 shows a hydraulic system to which a pressure compensating valve according to the present invention is applied. This hydraulic system is applied to, for example, a hydraulic excavator.

The hydraulic system includes a variable displacement hydraulic pump 1, a pilot hydraulic pump 2, a plurality of operation valves (directional control valves) 4 to which the discharge oil of the hydraulic pump 1 is supplied via an oil passage 3, and each operation valve. 4 are provided with a plurality of hydraulic cylinders 5 and the like.

The hydraulic cylinder 5 has a head side oil chamber having an oil passage 6a.
And the pressure compensation valve 7 is connected to the operation valve 4, and the bottom oil chamber is connected to the operation valve 4 via the oil passage 6b. Although a pressure compensating valve is actually interposed in the oil passage 6b, the pressure compensating valve is omitted in FIG. 1 to avoid complicating the drawing. .

The load pressure of the hydraulic cylinder 5 connected to the oil passages 6a and 6b acts on the oil passages 6a and 6b. Shuttle valve 8
Detects the highest load pressure among the load pressure P _L acts on the oil passages 6a, 6b. The detected maximum load pressure P _LS is applied to the pump 1, each pressure compensating valve 7, the unload valve 1 via the oil passage 9.
0 and act on the load pressure bleed valve 11.

The hydraulic pump 1 has a control unit 12.
The control unit 12 compares the discharge pressure P _P of the pump 1 with the maximum load pressure P _LS, and controls the pump 1 so that the discharge pressure P _P is always higher than the maximum load pressure P _LS by a set pressure difference. Control the displacement volume.

The unload valve 10 returns the oil discharged from the hydraulic pump 1 directly to the tank to put the hydraulic pump 1 in a no-load state. The unload valve 10 has a main valve 100 and an electromagnetic proportional pressure control valve 101, and operates as follows.

The main valve 100 is interposed between the discharge path 3 of the pump 1 and the tank. The main valve 100 returns the oil discharged from the pump 1 to the tank and communicating actuated when the difference between the discharge pressure P _P and the load pressure P _LS of the pump 1 (P _P -P _LS) is raised to the unload starting pressure.

[0046] Note that this unloading valve 10 is an electromagnetic proportional pressure control valve 101 generates the pilot pressure P ₅ reducing the pressure of the discharge pressure P ₄ of the pilot hydraulic pump 2. Meanwhile the main valve 100 is operated starting pressure depending on the pilot pressure P ₅ supplied from the input electromagnetic proportional pressure control valve 101 (unload starting pressure) changes.

Therefore, according to the unload valve 10, the unload start pressure can be set to a desired value by changing the control signal applied to the solenoid of the electromagnetic proportional pressure control valve 101.

When each of the operation valves 4 is operated to the neutral position, the differential pressure (P _P -P _LS ) rises to the unload start pressure, so that the discharge oil of the pump 1 is discharged through the unload valve 10. The pump 1 is returned to the tank, and as a result, the pump 1 is in a no-load state.

The load pressure bleed valve 11 is provided for adjusting the rising speed of the discharge pressure of the pump 1.
The load pressure bleed valve 11 has a variable throttle valve 110 and an electromagnetic proportional pressure control valve 111, and operates as follows.

When each of the operation valves 4 is operated to the neutral position, the pressure oil in the oil passage 9 flows out to the tank through the throttle 112, so that the maximum load pressure P acting on the volume control unit 12 of the pump 1 is obtained. _LS gradually decreases, and when the maximum load pressure P _LS decreases to zero, the volume of the pump 1 becomes minimum.

The variable throttle valve 110 is interposed between the oil passage 9 and the tank. Therefore, when one or both of the operation valves 4 are operated from the neutral position, the load pressure P _{LS of the} oil passage 9 is increased.
Rises, a part of the pressure oil in the oil passage 9 is
It will flow out into the tank via 10.

[0052] rate of increase in the discharge pressure P _P of the pump 1 is dependent on the rate of increase of hydraulic P _LS acting on the displacement control member 12. The rate at which the hydraulic pressure _PLS rises depends on the opening area between the inlet and outlet ports of the variable throttle valve 110, and increases as the flow resistance increases.

[0053] Therefore if the variable throttle valve 110 aperture amount appropriate settings, it is possible to start the cylinder 5 smoothly by preventing a rapid rise of the discharge pressure P _P of the pump 1. That is, it is possible to reduce the shock at the time of starting the work machine (the work machine such as a boom, an arm, and a bucket in a hydraulic shovel) driven by the cylinder 5.

It should be noted the electromagnetic proportional pressure control valve 111 of the load pressure bleed valve 11 outputs the pilot pressure P ₆ reducing the pressure of the discharge pressure P _P of the pilot hydraulic pump 2. On the other hand, the variable throttle valve 110 has an opening area between an inlet port and an outlet port whose control pressure P is given by the electromagnetic proportional pressure control valve 111.
It grows with the increase of ₆ .

[0055] Therefore, according to this load pressure Pulido valve 11, setting the rate of increase of the discharge pressure P _P of the pump 1 by changing the control signal applied to the solenoid of the electromagnetic proportional pressure control valve 111 to a desired size can do.

Hereinafter, the structure of the pressure compensating valve 7 according to the present invention will be described with reference to FIG. The pressure compensating valve 7 includes a compensating section 7A, a control pressure generating section 7B, and a pilot pressure supplying section 7C.

The compensator 7A has a main valve 20. The main valve 20 includes a first pressure receiving section 21, a second pressure receiving section 22, and a third pressure receiving section 23. The pressure P ₀ acting on the first pressure receiving portion 21 acts to increase the opening area between the inlet port 24 and the outlet port 25. The pressure P ₃ acting on the second pressure receiving part 22 and the pressure P ₁ acting on the third pressure receiving part 23 act together with the elastic force of the spring 26 to reduce the opening area.

The inlet port 24 is connected to the outlet port of the operation valve 4 shown in FIG. The pressure P ₀ at the inlet port 24 acts on the first pressure receiving portion 21 via the oil passage 27. The outlet port 25 is connected to the oil passage 6a via a load check valve 28.

The shuttle valve 29 controls the larger one of the load pressure P _L acting on the oil passage 6a and the maximum load pressure P _LS P _3.
And the pressure P ₃ is applied to the second pressure receiving portion 22 of the main valve 20.

The control pressure generator 7 B has a variable throttle valve 30. The variable throttle valve 30 operates so as to reduce the opening area between the inlet port 32 and the outlet port 33 by the elastic force of the spring 31. The opening area is increased by the pilot pressure P ₂ acting on the pressure receiving portion 34 and the elastic force of the spring 35.

The spring 35 is normally used only for the initial fine adjustment of the variable throttle valve 30.
Not essential. The reason why the tank port pressure is always applied to the spring 31 is to ensure that the variable throttle valve 30 operates more quickly.

The inlet port 32 of the variable throttle valve 30 is connected to the inlet port 24 of the main valve 20 via an oil passage 37 having a throttle 36. Inlet port 3 of variable throttle valve 30
The pressure P ₁ is the third pressure of the main valve 20 via the oil passage 38.
It acts on the pressure receiving part 23.

The outlet port 33 of the variable throttle valve 30 is connected to the oil passage 6a via an oil passage 40 having a check valve 39. Thus the pressure P ₁ acting on the third pressure receiving portion 23 of the main valve 20, the pressure P ₀ and the load pressure P _L and the aperture 36 and the variable throttle valve 3 in the inlet port 24 of the main valve 20
It is determined by the aperture amount of 0. When the variable throttle valve 30 is in the closed state, P ₁ = P ₀ .

The pilot pressure P ₂ is given as an output pressure of an electromagnetic proportional pressure control valve 50 provided in the pilot pressure supply section 7C. This electromagnetic proportional pressure control valve 50 is
The pressure oil discharged from the pilot hydraulic pump 2 shown in FIG. The pressure P _{4 of} this pressure oil
Is controlled by the pilot pressure P
_The pressure is reduced to ₂ . The pilot pressure P ₂ indicates a magnitude proportional to the amount of current supplied to the solenoid 53.

When the amount of current supplied to the solenoid 53 is zero, the outlet port 55 communicates with the tank port 56 by the elastic force of the spring 54 as shown in the figure. Therefore, the pressure P acting on the pressure receiving portion 34 of the variable throttle valve 30
₂ becomes zero. Accordingly, the variable throttle valve 30 is shut off between the inlet port 32 and the outlet port 33 by the elastic force of the spring 31 in which the tank port pressure acts on both sides. The discharge pressure of the pilot hydraulic pump 11 is kept constant by a constant-pressure means (not shown).

Next, specific structures of the operation valve 4 and the pressure compensating valve 7 will be described. As described above, in the hydraulic system of FIG. 1, the pressure compensating valve 7 is interposed not only in the oil passage 6a but also in the oil passage 6b.

FIG. 3 exemplifies the structure of the operation valve 4 for selectively supplying pressure oil to both the pressure compensating valves 7. The operation valve 4 has a structure in which a spool 61, a pair of left and right outlet ports 62, a pair of left and right pump ports 63, a pair of left and right actuator ports 64, and a pair of left and right tank ports 65 are provided on the body 60.

The spool 61 blocks all the ports 62 to 65 in the illustrated neutral state. When the spool 61 moves leftward from this neutral state, one of the outlet ports 62
And the pump port 63 communicate with each other, and the other actuator port 64 and the tank port 65 communicate with each other. When the spool 61 moves rightward from the neutral state, the other outlet port 62 and the pump port 63 communicate with each other, and the one actuator port 64 and the tank port 65 communicate with each other.

The main valve 20 provided in the compensating portion 7A of the pressure compensating valve 7 includes a valve portion 66A interposed between the outlet port 62 of the operation valve 4 and the actuator port 64,
6A.

FIG. 4 is an enlarged view of the pressure compensating valve 7. As shown in FIG. 4, the valve portion 66 has a hollow portion 68 opening at the left end, a hole 69 passing through the hollow portion 68 and opening on the outer peripheral surface, and a seat press-contacting a seat 70 formed on the body 60. And a surface 71.

The pressure receiving surfaces 66a and 66b of the valve portion 66 constitute the first pressure receiving portion 21 of the main valve 20 shown in FIG. 2, and the hole 69 of the valve portion 66 constitutes the outlet port 25 of the main valve 20. doing. The entire valve portion 66 has a function as the load check valve 39 shown in FIG.

The pushing portion 67 is located on the extension of the central axis of the valve portion 66 and slides left and right within a sleeve 72 solidified in the body 60, and slides left and right within the piston 73. And a spring 74 (see FIG. 2) interposed between the sleeve 72 and the slider 74.

Between the body 60 and the sleeve 72, there is formed an annular space 75 into which the pressure oil having the maximum one load pressure P _LS (see FIG. 2) is introduced. The pressurized oil of the maximum load pressure P _LS introduced into the annular space 75 is _{supplied to} the fine hole 76 provided in the sleeve 72, the annular groove 77 and the hole 7 provided in the piston 73.
8 flows into the stepped hole 80 in the slider 74 via the inlet port 79 provided in the slider 74 and acts on the right side of the ball 81 provided in the stepped hole 80.

[0074] On the other hand the hydraulic fluid actuator port 64 of the operation valve 4 shown in FIG. 3, i.e. the pressure oil of the load pressure P _L that flows through the oil passage 6a is through the inlet port 82 provided at the left end portion of the piston 73 To flow into the stepped hole 80 of the slider 74,
Acts on the left side of the ball 81.

The relationship between the pressures P _LS and P _L is P _LS > P _L
Ball 81 when the rolling to the left side position of the outlet port 84, the ball 81 when the P _LS <P _L to roll to the right side position of the outlet port 84. As shown in FIGS. 1 and 2, P _LS <P _L is a transitional state, and the pressure P _LS increases as P _LS = P _L.

The stepped hole 80 communicates with the pressure chamber 83 through an outlet port 84 and a concave groove 85 provided on the outer peripheral surface thereof. Maximum load pressure P when the thus P _LS> P _L
LS pressure oil is introduced into the pressure chamber 83, and P _LS <P _L
Pressurized oil of the load pressure P _L is introduced into the pressure chamber 83 when the.

As described above, the stepped hole 80 and the ball 81 have a function of detecting the higher one of the hydraulic pressures P _LS and P _L and guiding the detected pressure to the pressure chamber 83. The shuttle valve 29 shown in FIG. 2 includes the stepped hole 80 and the ball 81. Thus the pressure of the hydraulic fluid introduced into the pressure chamber 83 is a pressure P ₃ shown in FIG.

The pressure chamber 83 is a space surrounded by the inner surface of the sleeve 72, the right end surface of the piston 73, and the outer peripheral surface of the slider 74. The right end surface of the piston 73 has the second pressure receiving portion shown in FIG. It has a function as 22.

Next, the control pressure generator 7B will be described. The control pressure generating section 7B is located on the side of the compensating section 7A and includes the variable throttle valve 30 shown in FIG.

The body 87 of the variable throttle valve 30 is provided with a spool 88 for changing the flow resistance (amount of throttle) between the inlet port 32 and the outlet port 33 shown in FIG.

The spool 88 is given a downward force (in the direction of increasing the flow resistance) by the spring 31, and
An upward force (flow resistance decreasing direction) is applied by a spring 35 in a pressure chamber 90 formed between the adjusting screw 89 and the adjusting screw 89. The spool 8 facing the pressure chamber 90
8 constitutes the pressure receiving portion 34 shown in FIG.

A concave portion 91 provided on the left side of the body 87
Forms a pressure chamber 92 together with the right end face of the sleeve 72 and the right end face of the slider 74 of the compensating portion 7A. The right end face of the slider 74 facing the pressure chamber 92 constitutes the second pressure receiving portion 23 of the main valve 20 shown in FIG.

The inlet port 32 of the variable throttle valve 30 is connected to the outlet port 62 of the operation valve 4 via the oil passage 37 provided with the throttle 36, that is, the input port 24 of the main valve 20 shown in FIG.
And the pressure chamber 92 via the oil passage 38. The output port 33 is in communication with an actuator port 64 of the operation valve 4 via an oil passage 40 provided with a check valve 39 (see FIG. 2).

The pilot pressure generator 7C is provided above the body 87 of the control pressure generator 7B. The electromagnetic proportional pressure control valve 50 constituting the pilot pressure generating section 7C includes a spool 94 vertically disposed on the body 93, and a solenoid 53 for pushing the spool 94 against the spring 54.

In this electromagnetic proportional pressure control valve 50,
By lowering the spool 94 by the thrust of the solenoid 53, the flow resistance between the inlet port 52 and the outlet port 55 is reduced.

The outlet port 55 communicates with the pressure chamber 90 of the variable throttle valve 30 via an oil passage 95. The spool 94 is located on the axis of the spool 88 of the control pressure generator 7B.

Next, the operation of the pressure compensating valve 7 having the above configuration will be described with reference to FIG.

The pressure oil of pressure P ₀ flowing out of the outlet port 62 of the operation valve 4 acts on the surfaces 66a and 66b of the valve portion 66 constituting the pressure receiving portion 21 of the main valve 20 shown in FIG. Push the valve 66 to the right.

On the other hand, the load pressure P ₃ flowing into the pressure chamber 83
The pressure oil (pressure P _L or P _LS ) acts on the right end face of the piston 73 (the second pressure receiving portion 22 shown in FIG. 2) to push the valve portion 66 to the left and to control the pressure flowing into the pressure chamber 92. the right end surface of the pressure oil P ₁ is slider 74 (third pressure receiving portion shown in FIG. 2 2
Acting on 3), pushes the valve 66 to the left. The spring 26 also pushes the valve 66 to the left via the slider 74.

Thus, the pressure balance in the main valve 20 can be expressed by the following equation (1).

P ₀ × A ₀ = P ₁ × A ₁ + P ₃ (A ₀ −A ₁ ) + F ₀ (1) A ₀ > A ₁ where A _{0 is} the area of the surfaces 66 a and 66 b of the valve section 66. Sum A ₁ : Area of right end face of slider 74 A ₀ -A ₁ : Area of right end face of piston 73 F ₀ : Elastic force of spring 26

The pressure P _{1 in the} equation (1) is a control pressure for changing the pressure compensation characteristic of the pressure compensation valve 7. The control pressure P ₁ is controlled by the control pressure generator 7 shown in FIG.
Variable throttle valve 30 B is the pressure P ₀ in a state that closed. Therefore, when P ₀ is substituted for P ₁ in the equation (1), the relation of the following equation (2) is obtained.

P ₀ −P ₃ = F ₀ / (A ₀ −A ₁ ) (2)

As is apparent from this relationship, the pressure compensating valve 7 operates so that the pressure difference P ₀ -P ₃ becomes constant under the condition of P ₁ = P ₀ . That is, a pressure compensation operation is performed.

Therefore, when the two operation valves 4 shown in FIG. 1 are operated to operate the respective cylinders 5 in a combined manner, the inconvenience that pressure oil is concentratedly supplied only to the side of the cylinder 5 with a light load is avoided. You.

When the variable throttle valve 30 of the control pressure generator 7B is not closed, the pressure oil that has passed through the fixed throttle 36 flows to the cylinder 5 via the variable throttle valve 30 and the check valve 39. Therefore, the third pressure receiving portion 23 of the main valve 20 of the compensator 7A restricts the differential pressure between the pressures P ₀ and P _L
Variable throttle valve aperture ratio control pressure P ₁ divided by 30, i.e. the control pressure P ₁ that the pressure was reduced P ₀ will act as. In this state, the leftward moving force of the slider 74 shown in FIG. 4 is lower than when P ₁ = P ₀ .

Reducing the leftward moving force of the slider 74 is equivalent to reducing the elastic force F ₀ of the spring 26 in the above equation (2). Therefore, if the control pressure P ₁ is P
If it is lower than ₀ would differential pressure P ₀ -P ₃ are set lower than when P ₁ = P ₀ (change in pressure compensation characteristics). Here, even if the pressure compensation characteristic changes, the pressure difference P
The ability to maintain the ₀ -P ₃ constant is still maintained.

In the case of two or more cylinder circuits having different loads, under the condition that the operation amount of the operation valve 4 is constant, a large amount of pressure oil flows toward a lower load.

The control pressure P ₁ decreases as the amount of current supplied to the solenoid 53 of the electromagnetic proportional pressure control valve 50 increases. Therefore, for example, when a working machine of a construction machine (a boom, an arm, a bucket, etc. in the case of a hydraulic shovel) is driven by the cylinder 5, the pressure applied to the working mode of the working machine is controlled by controlling the amount of electricity supplied to the solenoid 53. Compensation characteristics can be set.

Of course, it is also possible to make the pressure compensation characteristics of the pressure compensating valves 7 for the plurality of cylinders 5 shown in FIG. 1 different from each other. Further, it is also possible to make the operation speed when the cylinder 5 expands and the operation speed when the cylinder 5 contracts by making the pressure compensation characteristics of the pressure compensating valves 7 relating to the single cylinder 5 shown in FIG. 3 different from each other.

Incidentally, in the pressure compensating valve 7, when the solenoid 53 of the electromagnetic proportional pressure control valve 50 is disconnected or the pilot pump 2 shown in FIG. the pilot pressure P ₂ does not act on the valve 39. That is, the variable throttle valve 39
Stops the aperture operation.

However, even if such a situation occurs, the pressure compensating function of the pressure compensating valve 7 is not impaired.
It is only in a completely compensated state.

That is, when the variable throttle valve 39 is closed, it becomes impossible to change the magnitude of the control pressure P ₁ to change the pressure compensation characteristic. However, since the control pressure P ₁ is set to P ₁ = P ₀ , the pressure compensation operation for keeping the pressure difference P ₀ −P ₃ shown in the above equation (2) constant is maintained.

FIG. 5 shows a second configuration example of the pressure compensating valve 7. The pressure compensating valve 7 connects the spring 54 of the electromagnetic proportional pressure control valve 50 to the variable throttle valve 30 of the control pressure generator 7B.
4 differs from the pressure compensating valve 7 in FIG.

[0105] In this pressure compensating valve 7, when the variable throttle valve based on the pilot pressure P ₂ supplied from the electromagnetic proportional pressure control valve (50) (30) spool (88) of the actuation, the actuation force It is mechanically fed back to the spool (94) of the electromagnetic proportional pressure control valve (50) via the spring 54. Therefore the variable throttle valve (30)
The dynamic characteristics (response) of the spool (88) are improved, and more accurate pressure compensation becomes possible.

FIG. 6 shows a third configuration example of the pressure compensating valve 7. The pressure compensating valve 7 has a body 1 in which the variable throttle valve 30 of the control pressure generating unit 7B and the electromagnetic proportional pressure control valve 50 are common.
The solenoid 53 of the electromagnetic proportional pressure control valve 50 is provided outside the body 100. Therefore, the control pressure generating portion 7B of the pressure compensating valve 7 forms a flange portion 88a having a tapered peripheral surface on the spool 88 of the variable throttle valve 30, while making the configuration compact and reducing the number of parts. The flange portion 88a is interposed between the inlet port 32 and the outlet port 33 of the variable throttle valve 30.

[0107] pressure when pressure oil of the pressure P _L According to this structure flows into the outlet port 33 of the variable throttle valve 30 via the oil passage 40, the upper surface of the flange portion 88a is by this pressure oil Is done. Accordingly, the spool 88 moves down, and the tapered peripheral surface of the flange portion 88a is pressed against the seating surface of the body 100. As a result, the space between the inlet port 32 and the outlet port 33 is shut off.

As described above, the spool 88 has a function as a check valve for preventing the pressure oil of the pressure P _{L from} flowing to the inlet port 32 side. Therefore, this pressure compensating valve 7
Is the check valve 3 shown in FIGS.
It is not necessary to provide 9, which makes the processing of the body 100 easy.

FIG. 7 shows a fourth configuration example of the pressure compensating valve 7. The pressure compensating valve 7 is connected to a mounting block 101 fixed to the upper surface of the body 87 of the control pressure generating section 7B by a joint 1.
02, and the pressure chamber 90 of the variable throttle valve 30 of the control pressure generating unit 7B is connected to the outlet port 55 of the electromagnetic proportional pressure control valve 50 via the joint 102 and the pipe 95.

In the pressure compensating valve 7, the pilot pressure P ₂ output from the electromagnetic proportional pressure control valve 50 or the pilot pressure output from the manual pilot valve (not shown) is applied to the control pressure generating section via the joint 102. 7
B can be applied to the variable throttle valve 30. Therefore, the pressure compensating valve 7 is suitable for use when the electromagnetic proportional pressure control valve 50 and the pilot valve have to be provided separately from the control pressure generating section 7B due to the limitation of the arrangement space and the like.

The variable throttle valve 30 of the control pressure generator 7B of the pressure compensating valve 7 has the same configuration as the variable throttle valve 30 of the pressure compensating valve 7 shown in FIG.

FIG. 8 shows a fifth configuration example of the pressure compensating valve 7. The pressure compensating valve 7 has a joint 104 attached to the outside of the body 103 of the control pressure generating section 7B, and the pressure chamber 9 provided in the variable throttle valve 30 of the control pressure generating section 7B.
0 is connected to the electromagnetic proportional pressure control valve 50 or a manual pilot valve (not shown) through the joint 104 and the pipe 95.

This pressure compensating valve 7 is an electromagnetic proportional pressure control valve 5
Zero or a pilot valve can be provided at a distance from the control pressure generator 7B. Since the pressure compensating valve 7 is provided with the joint 104 on the body 103 of the control pressure generating section 7B, it is possible to reduce the size and the number of parts.

The variable throttle valve 30 of the control pressure generator 7B
Has the same configuration as the variable throttle valve 30 of the pressure compensating valve 7 shown in FIG. Therefore, this pressure compensating valve 7 is
The body 1 of the control pressure generator 7B is similar to the pressure compensating valve 7 of FIG.
There is no need to provide a check valve at 03.

FIG. 9 shows a sixth configuration example of the pressure compensating valve 7. The pressure compensating valve 7 includes only a compensating section 7A and a control pressure generating section 7B. The compensator 7A has the same configuration as the compensator 7A shown in FIG.

The control pressure generator 7B has a variable throttle valve 30 having a structure for manually changing the size of the throttle. The variable throttle valve 30 has a vertical hole 106 formed in the body 105, and a poppet valve type spool 107 is inserted into the vertical hole 106. The upper and lower parts of the vertical hole 110 are spool 107
Can be communicated and blocked by vertical movement of

The upper portion of the vertical hole 109 communicates with the actuator port 64 of the operation valve 4 via the oil passage 40.
The lower part of the vertical hole 106 is provided with an oil passage 37 having the throttle 36.
Through the oil passage 38 and the pressure chamber 92 through the oil passage 38.

An adjusting screw 108 is screwed above the vertical hole 106, and a spring 109 having a low elastic force is interposed between the adjusting screw 108 and the spool 107.

In the variable throttle valve 30 configured as described above, the pressure oil of the pressure P ₀ discharged from the outlet port 62 of the operation valve 4 flows into the lower part of the vertical hole 106 via the oil passage 37. .

Accordingly, the spool 107 is pushed up, and a part of the pressure oil at the pressure P ₀ flows into the oil passage 40 while being throttled by the spool 107. The pressure P in the pressure chamber 92
₁ is set according to the amount of pressure oil flowing into the oil passage 40, that is, according to the throttle amount of the spool 107.

The upward movement stroke of the spool 107 that defines the throttle amount of the spool 107 is adjusted by the adjusting screw 10.
8 can be adjusted by rotating it manually. Therefore, the pressure compensating valve 7 is
It is possible to change the pressure P ₁ by rotating, that is, to change the pressure compensation characteristic.

Since the spool 107 is of a poppet valve type, when pressure oil flows into the oil passage 40 from the cylinder 5 side, the spool 107 is pushed downward to shut off the upper and lower portions of the vertical hole 106. . In other words, this spool 1
07 has a function as a check valve.

Therefore, according to the pressure compensating valve 7, there is no need to provide the check valve 39 shown in FIG. 4 on the body 105, and therefore, the processing of the body 105 is facilitated.

FIG. 10 shows a seventh configuration example of the pressure compensating valve 7. This pressure compensating valve 7 is different from the pressure compensating valve 7 shown in FIG.

That is, the main valve 20 of the compensator 7A shown in FIG. 10 has the spool S in which the valve 66 and the pusher 67 shown in FIG. 4 are integrated.

In this pressure compensating valve 7, since the pressure oil of the maximum load pressure P _LS introduced into the annular space 75 flows directly into the pressure chamber 83 through the hole 112 provided in the sleeve 72, the pressure chamber pressure P ₃ of the 83 becomes the maximum load pressure P _LS.

The spool S is connected to the communication hole 11 along the central axis.
3 by which the outlet port 6 of said operating valve 4
2 and the pressure chamber 92 are communicated. Therefore, the pressure oil of the pressure P ₀ flowing out from the outlet port 62 of the operation valve 4 is:
It will flow into the pressure chamber 92 via the communication hole 113. That is, the communication hole 113 has a function as the oil passage 37 in FIG.

A fixed throttle 113a corresponding to the fixed throttle 36 shown in FIG. 4 is formed at the end of the communication hole 113 on the pressure chamber 92 side.

In the pressure compensating valve 7 having the above structure, FIG.
It is not necessary to provide the oil passage 37 shown in FIG. 4 in the body 60 of the compensating unit 7A, and it is not necessary to provide the throttle 36 shown in FIG. Therefore, the processing of the bodies 60 and 87 can be facilitated.

The variable throttle valve 30 of the control pressure generator 7B
Has a configuration similar to that of the variable throttle valve 30 shown in FIG.

FIG. 11 shows an eighth configuration example of the pressure compensating valve 7. This pressure compensating valve 7 has a compensating section 7A having a structure shown in FIG.
0, and the configuration of the control pressure generator 7B and the pilot pressure generator 7C is the same as that of the pressure compensator 7 shown in FIG.

Therefore, according to the pressure compensating valve 7, the effect of the pressure compensating valve 7 shown in FIG. 10 that the processing of the body 60 and the body 100 can be facilitated can be obtained. FIG. 6 that the reduction of the size and the ease of processing the body 100 can be achieved.
The effect of the pressure compensating valve 7 shown in FIG.

FIG. 12 shows a ninth configuration example of the pressure compensating valve 7. This pressure compensating valve 7 is the same as the pressure compensating valve 7 shown in FIG. 10 in the configuration of the compensating unit 7A, and is different from the pressure compensating valve 7 shown in FIG. Are identical.

According to the pressure compensating valve 7, the body 60,
The effect of the pressure compensating valve 7 shown in FIG. 10 and the electromagnetic proportional pressure control valve 5 shown in FIG.
0 can be provided at a distance from the control pressure generating section 7B, and both effects of the pressure compensating valve 7 shown in FIG. 7 can be obtained.

FIG. 13 shows a tenth configuration example of the pressure compensating valve 7. The pressure compensating valve 7 has the same configuration of the compensating unit 7A as the pressure compensating valve 7 shown in FIG. 10, and the configuration of the control pressure generating unit 7B and the mounting position of the joint 104 are the same as those of the pressure compensating valve 7 shown in FIG. Are identical.

According to the pressure compensating valve 7, the effect of the pressure compensating valve 7 shown in FIG. 10 that the processing of the body 60 and the body 100 can be facilitated is obtained, and the electromagnetic proportional pressure control valve is provided. 50 is the control pressure generator 7B
The effect of the pressure compensating valve 7 shown in FIG.

The variable throttle valve 30 of the control pressure generator 7B
Has the same configuration as the variable throttle valve 30 of the pressure compensating valve 7 shown in FIG. Therefore, the effect of the pressure compensating valve 7 shown in FIG. 6 that the check valve does not need to be provided in the body 103 of the control pressure generating unit 7B is also obtained.

FIG. 14 shows an eleventh configuration example of the pressure compensating valve 7. This pressure compensating valve 7 has the same configuration of the compensating unit 7A as the pressure compensating valve 7 shown in FIG. 10, and the configuration of the control pressure generating unit 7B is the same as the pressure compensating valve 7 shown in FIG.

According to the pressure compensating valve 7, the body 60,
FIG. 10 that the processing of 105 can be facilitated.
The effect of the pressure compensating valve 7 shown in FIG. Further, the pressure compensating valve 7 shown in FIG. 9 is capable of manually adjusting the amount of drawing and facilitating the processing of the body 105.
Also has the effect.

FIG. 15 shows a twelfth configuration example of the pressure compensating valve 7. This pressure compensating valve 7 is different from the pressure compensating valve 7A shown in FIG.

The spool S of the main valve 20 of the compensating section 7A shown in FIG. 15 includes a piston 116 in which the valve section 66 and the piston 73 shown in FIG. 4 are integrated, and a slider 117 provided in the piston 116. It has.

The piston 116 and the slider 117 have communication holes 118 and 119, respectively, along the central axis. One end of the communication hole 119 of the slider 117 communicates with the communication hole 118 of the piston 116 via the check valve 120, and the other end communicates with the pressure chamber 92 via the throttle 119a corresponding to the throttle 36 of FIG. ing.

According to the pressure compensating valve 7 having the above structure, the pressure oil of the pressure P ₀ supplied from the outlet port 62 is supplied to the communication hole 1.
18, a check valve 120, a slit 121 formed around the check valve 120, a port 122 penetrating the peripheral wall of the slider 117, a communication hole 119, and a throttle 119a.
Flows into the pressure chamber 92 via That is, the communication hole 118,
Reference numeral 119 has a function as the oil passage 37 in FIG.

Therefore, it is not necessary to provide the oil passage 37 shown in FIG. 4 in the body 60 of the compensating section 7A, and it is not necessary to provide the throttle 36 shown in FIG. 4 in the body 87 of the control pressure generating section 7B. Therefore, the processing of the bodies 60 and 87 can be facilitated.

On the other hand, when the pressure P _L of the pressure oil at the actuator port 64 becomes higher than the pressure P ₀ of the pressure oil at the outlet port 62, the check valve 120 is closed. Therefore, the check valve 120 prevents the pressure oil in the actuator port 64 from flowing into the outlet port 62.

Thus, the check valve 120 has the same function as the check valve 39 shown in FIG. Therefore, according to the pressure compensating valve 7, the check valve 39 shown in FIG.
Need not be provided on the body 87 of the control pressure generating section 7B, and this contributes to facilitation of processing of the body 87.

In each pressure compensating valve 7 described above, the oil passage 40 connected to the outlet port 33 of the variable throttle valve 30 is
The actuator port 64 of the operation valve 4 shown in FIG.
a), the oil passage 40 is connected to the tank port 6
5 may be connected.

[Brief description of the drawings]

FIG. 1 is a hydraulic circuit diagram of a hydraulic system to which a pressure compensation valve according to the present invention is applied.

FIG. 2 is a hydraulic circuit diagram showing a configuration of a pressure compensating valve.

FIG. 3 is a vertical cross-sectional view showing a connection mode of a pressure compensating valve and an operation valve.

FIG. 4 is a longitudinal sectional view showing a configuration of a pressure compensating valve.

FIG. 5 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 6 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 7 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 8 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 9 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 10 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 11 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 12 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 13 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 14 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 15 is a longitudinal sectional view showing another configuration of the pressure compensating valve.

FIG. 16 is a hydraulic circuit diagram showing a configuration of a conventional hydraulic device including a pressure compensating valve.

FIG. 17 is a graph showing a relationship between a differential pressure and a control force.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Variable displacement hydraulic pump, 2 ... Pilot hydraulic pump, 3 ... Oil passage, 4 ... Operating valve, 5 ... Hydraulic cylinder, 6a,
6b: oil passage, 7: pressure compensation valve, 7A: guarantee unit, 7B: control pressure generation unit, 7C: pilot pressure generation unit, 8: shuttle valve, 9: oil passage, 10: unload valve, 11: load pressure Bleed valve, 20 ... Variable throttle valve, 21, 22, 23 ... Pressure receiving part, 24 ... Inlet port, 25 ... Outlet port, 26 ... Spring, 27 ... Oil passage, 28 ... Load check valve, 30
... variable throttle valve, 31 ... spring, 32 ... inlet port,
33 ... outlet port, 34 ... pressure receiving part, 36 ... throttle, 37,
38, 40 ... oil passage, 50 ... electromagnetic proportional pressure control valve, 52 ...
Inlet port, 53 ... solenoid, 54 ... spring, 5
5 outlet port, 60 body, 62 output port, 6
3 ... Pump port, 64 ... Actuator port, 65
... tank port, 66 ... valve part, 67 ... pushing part, 68 ... hollow part, 69 ... hole, 73 ... piston, 74 ... slider, 88
... Spool, 90 ... Pressure chamber, 92 ... Pressure chamber, 94 ... Spool, 102,103 ... Coupling, 107 ... Spool, 11
DESCRIPTION OF SYMBOLS 1 ... Spool, 113 ... Communication hole, 113a ... Restriction, 11
5 Spool, 118, 119 communication hole, 119a communication hole, 120 check valve, S spool.

Claims (9)

[Claims] An operation for increasing an opening area between an inlet port (24) and an outlet port (25) by a pressure applied to a first pressure receiving portion (21) and a second pressure receiving portion (22). The pressure acting on the pressure port (24) and the pressure acting on the third pressure receiving part (23) operate to reduce the opening area, and the pressure (P
₀ ) to the first pressure receiving portion (21), and the pressure (P ₃ ) of the load (5) driven by the pressure oil flowing out from the outlet port (25) to the second pressure receiving portion (22). A main valve (20) to be actuated, and a control pressure generation to act on the third pressure receiving part (23) a control pressure (P ₁ ) that reduces the pressure (P ₀ ) of the inlet port (24). Means (7B). 2. The control pressure generating means (7B) includes a fixed throttle (36) and a variable throttle valve (3) connected in series with each other.
0), and the pressure oil from the pump side is made to flow through the fixed throttle (36) and the variable throttle valve (30) to extract the hydraulic pressure generated at the connection point thereof as the control pressure (P ₁ ). The pressure compensating valve according to claim 1, wherein the pressure compensating valve is provided. 3. The variable throttle valve (30) operates in a direction to reduce an opening area between an inlet port (32) and an outlet port (33) by an elastic force of a spring (24),
_{3. The device} according to claim ₂ , wherein a pilot pressure (P ₂ ) supplied from a pilot pressure supply means (7 C) operates to increase an opening area between the inlet port (32) and the outlet port (33). 4. The pressure compensating valve as described. 4. A main valve (20) and a variable throttle valve (30) integrally connected to the main valve (20), wherein the main valve (20) has an inlet port (24) and an outlet port (30). 25) a spool (S) provided with a valve portion (66) for shutting off communication between the portions and a pressing portion (67) for applying a pressing force to the valve portion (66); A first pressure receiving portion (21) for receiving a hydraulic pressure for operating the (S) in the communication direction is formed, and a hydraulic pressure for operating the spool (S) in the shutoff direction is received by the pushing portion (67). Second
And a third pressure receiving portion (23), and communicates between an inlet port (24) of the main valve (20) and an inlet port (32) of the variable throttle valve (30). A first oil passage (37, 113, 118, 119); a fixed throttle (36, 113a, 119a) interposed in the first oil passage (37, 113, 118, 119); A second oil passage (40) for communicating the outlet port (25) of the valve (20) with the outlet port (33) of the variable throttle valve (30); and a second oil passage (40) interposed between the second oil passage (40). A check valve (39, 120) for preventing a flow of pressure oil from an outlet port (33) of the variable throttle valve (30) to an outlet port (25) of the main valve (20); 30) the inlet port (32) and the third pressure receiving section (23) are connected. The third oil passage to (38)
And a pressure compensating valve, wherein a load pressure (P ₃ ) is applied to the second pressure receiving portion (22). 5. An outlet port (25) of said main valve (20).
The load pressure (P _L ) of the load (5) driven by the pressure oil flowing out of the second pressure receiving section (22) is compared with the load pressure of another load, and the larger load pressure (P ₃ ) is applied to the second pressure receiving section (22). 5) The pressure compensating valve according to claim 4, wherein a load pressure comparing means (29) acting on the main valve (20) is incorporated in the main valve (20). 6. The spool (8) of the variable throttle valve (30).
8), an electromagnetic proportional pressure control valve (50) for applying a variable pilot pressure is added.
0) positions the spool (94) on the axis of the spool (88) of the variable throttle valve (30), and applies the operating force of the spool (88) to the spool (94) via the spring (54). The pressure compensating valve according to any one of claims 4 to 5, wherein the pressure compensating valve is arranged to provide a mechanical feedback. 7. The pressure compensating valve according to claim 4, wherein the valve portion (66) and the pushing portion (67) of the spool (S) are integrated. 8. The first oil passage (113, 118, 119) is formed in the spool (S), and the fixed throttle (113) is formed in the first oil passage (113, 118, 119).
a, 119a) is provided. The pressure compensating valve according to claim 7, wherein the pressure compensating valve is provided. 9. The pressure compensating valve according to claim 8, wherein the check valve (120) is provided in the first oil passage (118, 119).