JP2889317B2 - Pressure compensation valve - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial Application Field] The present invention relates to a pressure upstream and downstream of a flow control valve constituting a directional control valve provided in a hydraulic machine having a load sensing system such as a hydraulic shovel. And a pressure compensating valve for controlling a differential pressure.

[Conventional technology]

FIG. 4 is a longitudinal sectional view showing the structure of a directional switching valve having a conventional pressure compensating valve of this type, and FIG. 5 is a load sensing system provided with the directional switching valve having the pressure compensating valve shown in FIG. It is a circuit diagram which shows the hydraulic drive device of the civil engineering / construction machine with which it is equipped.

The hydraulic drive device shown in FIG. 5 includes a variable displacement hydraulic pump 1 that is a pressure oil supply source, a pump flow control device 2 for controlling the displacement of the variable displacement hydraulic pump 1, and a pump 1
A relief valve 3 for regulating the pressure of the hydraulic pressure discharged from the
An actuator driven by pressure oil discharged from the pump 1, for example, a swing motor 4, and another actuator not shown, and a direction switch for controlling a flow of the pressure oil supplied from the pump 1 to the swing motor 4. A valve 5 is provided.

As shown in FIG. 4, the direction switching 5 includes a block body 6 forming a main body, a flow control valve 8 having a spool 7 sliding in the block body 6, and an upstream of the flow control valve 8. provided on the side, the meter-in variable throttle block 14a of the flow control valve in the 8, the load pressure P _L upstream of the downstream of the first pressure receiving portion 22 in which the pressure Pz is given in 14b provided 2
Of having a pressure receiving portion 24, these upstream pressures Pz and downstream differential pressure of the load pressure P _L, that is controlled to be the differential pressure [Delta] P 'which is set by the spring 20 to the differential pressure Pz-P _L the pressure with the main valve body 9a which is slidably disposed block body 6 for supplying despite variations fluctuations and load pressure P _L of the pump pressure Pd during the combined operation, a substantially constant flow rate to the swing motor 4 It has a compensating valve 9 and a shuttle valve 10 provided downstream of the flow control valve 8.

The block body 6 has two hydraulic supply passages 11a and 11b connected to the pump 1, and these hydraulic supply passages 11a and 11b.
Load passages 12a and 12b that can communicate with the swing motor 4 and tank passages 13a and 13b that can communicate with the load passages 12a and 12b are provided. The spool 7 has a hydraulic supply passage 11a and a load passage 12a.
Or the metered-in variable throttle portions 14a and 14b, which connect the pressure oil supply passage 11b and the load passage 12b, and open by an amount corresponding to the stroke of the spool 7, respectively.
These variable throttle portion 14a, is provided downstream of the 14b, the detection ports for detecting a load pressure P _L of the swing motor 4 shown in FIG. 5
15a, 15b, passages 16a, 16b connected to these detection ports 15a, 15b, passages connected to these passages 16a, 16b
17a, 17b, and a passage 18 that can communicate with these passages 17a, 17b. Above passages 16a, 16b, 17a, 17b, 18, together with the passage 19 which is communicated to the shuttle valve 10 shown in FIG. 5, the detection port 15a, the load pressure P _L of the swing motor 4 detected by 15b as a control pressure A transmission passage for transmission to one drive unit of the pump flow control device 2 shown in FIG. 5 is configured.
A pump pressure Pd is guided to the other drive unit of the pump flow control device 2, and the pump flow control device 2
The pressure differential pressure Pd and the load pressure P _L is set in advance, i.e., to control the displacement volume of the pump 1 so as to balance the pressure ΔP corresponding to the force of the spring 2a. That is, the pump 1 is controlled so as to always maintain the pump pressure Pd = the load pressure P _L + ΔP.
To control the flow rate. Then, the detection ports 15a, 1
5b is open to the tank passages 13a, 13b in an operating state near the neutral position of the spool 7, and the passages 16a, 16b, 17a, 1
A discharge passage connecting the transmission passages including 7b and 18 to the tank passages 13a and 13b is configured.

In the direction switching valve 5 configured as described above, the switching timing for the spool stroke of the flow control valve 8 such as the variable throttle portions 14a and 14b and the detection ports 15a and 15b is adjusted by controlling the spool 7 with the intention of driving the swing motor 4 independently. Assuming that the spool 7 is moved rightward in FIG. 4 from neutral, a description will be given with reference to a characteristic diagram showing the relationship between the stroke of the spool 7 and the opening area shown in FIG. In FIG. 6, a characteristic line 20a indicates an opening area between the detection port 15a and the tank passage 13a, and a characteristic line 20b indicates an opening area between the passage 18 and the passage 17b, that is, the tank passage 13b. 20c indicates an opening area between the detection port 15a and the load passage 12a, and a characteristic line 20d indicates a variable throttle unit.
14a shows an opening area. When the spool 7 of the flow control valve 8 moves rightward from the state shown in FIG. 4, the characteristic line shown in FIG.
As shown by 20a, first, the detection port 15a and the tank passage 13a
Is cut off. At this time, as shown by the characteristic line 20b, the passage 18 communicates with the tank passage 13b via the passages 17b and 16b and the detection port 15b. Further, when the spool 7 moves rightward, the gap between the passage 18 and the passage 17b is cut off, and the transmission passage including the passage 18 and the like and the tank passage 13
The discharge passage formed so as to communicate with b disappears. When the spool 7 further moves to the right, the characteristic line 20
As shown by c, and the opening detection port 15a is on the load passage 12a, the load pressure P _L of the swing motor 4 shown in Fig. 5, the detection port 15a, passage 16a, 17a, 18, the shuttle valve 10, the fifth It is transmitted to one drive unit of the pump flow control device 2 via a passage 19 shown in the figure. Thereby, the discharge pressure Pd of the pump 1
Rises to a pressure such that Pd = P _L + ΔP. Then, when the spool 7 further moves from the above state, the meter-in variable throttle portion 14a opens as shown by the characteristic line 20d in FIG. 6, and the hydraulic pump 1 shown in FIG. The supplied pressure oil is supplied to the pressure oil supply passage 11a shown in FIG.
The swing motor 4 is guided to the swing motor 4 via the variable throttle unit 14a and the load passage 12a, and the swing motor 4 is operated to drive a swing body (not shown). At this time, in the variable throttle unit 14a, the pressure compensation is not performed because the opening amount of the variable throttle unit 14 is ΔP [Since the pressure compensating valve 9 is driven independently by the swing motor 4, ΔP = ｛(P _z = Pd) P _Lな る. ]
Assuming that C is a constant and A is the opening area of the variable throttle unit 14a, the flow rate Q supplied to the swing motor 4 is Becomes The pressure compensating valve 9 does not function when the rotating motor 4 is driven alone, but when the rotating motor 4 and other actuators (not shown) are combined, the pressure compensating valve 9 is variable in order to realize good combined driving of these actuators. throttle portion 14a, is controlled to be constant at the set differential pressure ΔP before and after differential pressure Pz-P _L spring 20 14b '. In this case, when the pump pressure Pd increases with the combined operation, the flow rate passing through the meter-in variable throttle sections 14a and 14b increases, and the variable throttle section
The upstream pressure Pz of 14a, 14b is introduced into the first oil chamber 23 through the passage 26, and is added to the first pressure receiving portion 22. The variable throttle portion 14a, the port pressure is the downstream pressure 14b, i.e., the load pressure P _L is the passage 15a, 16a, is introduced 17a from the port unit 30 through the inside second oil chamber 25, the second It is added to the pressure receiving section 24. Here, the downward movement of the main valve body 9a acting on the main valve body 9a in FIG. 4 restricts the space between the pump passage 6a and the supply passages 11a and 11b by the throttle valve 26, and the pressure Δ set by the spring 20
P ′ is balanced against P ′ so that Pz−P _L = ΔP ′,
The flow rate is determined.

[Problems to be solved by the invention]

By the way, in the conventional pressure compensating valve 9 described above,
Statically, as described above, the swing motor 4 operates in a balanced manner in accordance with the differential pressure ΔP when driven alone and in the combined drive, according to the differential pressure ΔP ′. Pressure compensation control is likely to be unstable due to the influence of 20 forces or fluid flow force. Therefore, in consideration of this dynamic stability, it is necessary to provide a restricting portion or the like for restricting the flow rate in the fluid passage. That is, the pressure Pz upstream of the supply passages 11a and 11b shown in FIG.
A throttle valve is provided in the passage 27 leading to the second oil chamber 25.
Communicating, it is necessary to increase the damping to or a throttle portion is provided in the port portions 30 is a passage of the load pressure P _L to. However, such a constricted portion is provided by the pressure compensating valve 9
The opening resistance of the main valve element 9a, ie, the upward movement of FIG. 4, and the closing operation of the main valve element 9a, ie, the downward movement of FIG. For example, when the flow control valve 8 is neutral, the pump pressure Pd is introduced into the first oil chamber 23, and the second oil chamber 25 drops to the tank pressure.
9a is located to the lower limit of FIG. 4 against the force of the spring 20, and the throttle section 26 is in the fully closed state, but assuming the throttle amount considering simple dynamic stability as described above. In this case, the operation in the opening direction of the main valve body 9a, that is, the opening operation of the throttle portion 26 when the flow control valve 8 is switched from neutral to drive the swing motor 4 is delayed, so that the switching operation of the flow control valve 8 is performed. The responsiveness of the swing motor 4 to the swing motor deteriorates, and the operability of an actuator such as the swing motor is deteriorated, and the efficiency of work performed via the actuator is reduced.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-described circumstances in the related art, and has as its object to improve the response of an actuator to switching from a neutral state of a flow control valve and to improve the dynamic stability when the actuator is driven. And to provide a pressure compensating valve that can realize the following.

[Means for solving the problem]

In order to achieve this object, the present invention has a main valve body slidably provided in a block body, and cuts off a pressure oil supply passage provided in the block body and a load passage connected to an actuator. Alternatively, it communicates with the opening amount according to the operation amount, and controls the differential pressure between the upstream pressure of the meter-in side variable throttle portion connecting the pressure oil supply passage and the load passage and the downstream load pressure of the meter-in variable throttle portion. A first pressure receiving portion provided at one end of the main valve body and facing the first oil chamber into which the upstream pressure is introduced, and provided at the other end of the main valve body. And a second pressure receiving portion facing the second oil chamber into which the downstream load pressure is introduced. The first pressure receiving portion and the second pressure receiving portion are provided with the upstream pressure and the downstream load. Pressure compensating valve, wherein the pressure and the pressure are given to oppose each other, Forming a seat surface to the port portion of at least one of the oil chambers of the first oil chamber and second oil chamber,
Another valve body which can be joined to the seat surface and can be interlocked with the main valve body, and a flow rate limiting means capable of limiting a flow rate of the pressure oil flowing between the at least one oil chamber and the port are provided. It has a configuration.

[Action]

Since the present invention is configured as described above, when the flow control valve is neutral, the pressure upstream of the meter-in variable throttle section, that is, the pump pressure Pd is the first pressure receiving pressure of the main valve body facing the first oil chamber. The second oil chamber is provided with a tank pressure through a flow restricting means provided between the inside of the second oil chamber and the port portion, and the main valve body causes a pressure between the pump passage and the pressure oil supply passage. The variable throttle section of the meter-in is closed, and the pressure oil supply passage and the load passage connected to the actuator are cut off, and another valve element is interlocked with the closing operation of the main valve element. The port is closed by contacting the sheet surface of the port.

When such switching the flow control valve from the neutral state, as the other valve body variable throttle opening at the same time as the load pressure P _L of the meter is given to another valve body away from the seat surface of the port portion The main valve body moves in the opening direction in conjunction with the other valve body. That is, almost simultaneously with the switching operation of the flow control valve from neutral, the pump passage and the pressure oil supply passage are opened, and the pressure oil supply passage and the load passage communicate with each other via the pump passage, the pressure oil supply passage, and the load passage. As a result, pressure oil is supplied to the actuator to drive the actuator, and the responsiveness of the actuator to switching from the neutral state to the flow control valve can be improved.

The static stability during the driving of the actuator is determined by the force of the pressure upstream of the meter-in variable throttle provided to the first pressure-receiving part of the main valve body and the second force of the main valve body. The pressure is maintained in the same manner as in the prior art so that the pressure due to the pressure downstream of the variable throttle portion of the meter-in provided to the pressure receiving portion becomes a predetermined value, and the dynamic stability is determined by the movement of the main valve body. As a result, the flow of the pressure oil between the inside of the second oil chamber and the port portion is restricted by the flow restricting means, so that unnecessary movement of the main valve body is restricted, and the stability thereof can be improved. .

〔Example〕

Hereinafter, an embodiment of the pressure compensating valve of the present invention will be described with reference to the drawings.

FIG. 1 is a longitudinal sectional view showing a configuration of a directional switching valve having a first embodiment of the present invention, and FIG. 2 is equipped with a load sensing system provided with a directional switching valve having a pressure compensating valve shown in FIG. 1 is a circuit diagram showing a hydraulic drive device of a civil engineering / construction machine.

The directional control valve 5 shown in FIG. 1 is drawn so as to correspond to the directional control valve shown in FIG. 4, and has a block body 6 equivalent to the directional control valve shown in FIG. A flow control valve 8 having a slidably provided spool 7,
Load passages 12a and 1a that can communicate with the pressure oil supply passages 11a and 11b via a pump passage 6a, pressure oil supply passages 11a and 11b that can communicate with the pump passage 6a, and meter-in variable throttle portions 14a and 14b.
2b, tank passages 13a and 13b, detection port
15a, 15b, passages 16a, 16b, 17a, 17b, a shuttle valve 10, and a passage 18 are provided.

Also, the hydraulic circuit shown in FIG. 2 is drawn corresponding to the hydraulic circuit shown in FIG. 5 described above, and has the same variable displacement hydraulic pump 1, pump pressure Pd and load pressure as those shown in FIG.
A pump flow control device 2 for controlling the displacement of the variable displacement hydraulic pump 1 so as to have a flow rate corresponding to a pressure difference ΔP from the pressure P _L , a relief valve 3, a swing motor 4 exemplified as an actuator, a tank 13, and a load pressure A passage 19 is provided for guiding P _L to the pump flow control device 2.

The pressure compensating valve 9 of this embodiment is similar to the pressure compensating valve 9 shown in FIG.
And a first pressure receiving portion 22 that faces the first oil chamber 23 and receives the upstream pressure Pz.
When, and a main valve body 9a having a second pressure receiving portion 24 which faces the second oil pressure 25, subjected to the load pressure P _L.

The second pressure receiving portion 24 of the pressure compensating valve 9 is formed in a cylindrical shape, and another valve element 40 is provided inside the cylindrical portion.
Is slidably provided. The port portion 30 has a seat surface formed thereon, and the above-mentioned another valve element 40 can be joined to the seat surface. The other valve body 40 is provided with flow rate limiting means capable of limiting the flow rate of the pressure oil flowing through the second oil chamber 25 and the port 30, for example, a throttle section 41.
A passage 40a that communicates between the portion of the second oil chamber 25 located inside the second oil chamber 25 and the portion of the second oil chamber 25 located outside the other valve body 40
Is provided. Also, the inner portion of another valve body 40 and the main valve body 9a
Between the second pressure receiving portion 24 and another port 40
A spring 20a for urging the seat surface of the
The main valve element 9a and another valve element 40 can communicate with each other via 0a.

In this embodiment, the static movement is the same as that of the prior art shown in FIGS. 4, 5 and 6, and when the swing motor 4 is driven independently, the pressure compensating valve 9 is turned off. Due to the failure, the differential pressure Pz across the flow control valve 8
The displacement of the variable displacement hydraulic pump 1 is controlled such that the force due to −P _L, ie, Pd−P _L (= ΔP), and the force of the spring 2a of the pump flow control device 2 are balanced. The flow rate corresponding to the pressure difference ΔP is supplied to the swing motor 4 and the swing motor 4 is driven. When the swing motor 4 and an actuator not shown are combinedly driven, the pressure compensating valve 9 functions and the flow control valve 8 differential pressure Pz-P _L =
The force due to ΔP ′ and the force of the spring 20a are controlled so as to be balanced, and a flow rate corresponding to the differential pressure ΔP ′ is supplied to the swing motor 4 to realize a combined drive of the swing motor 4 and another actuator. be able to.

When the flow control valve 8 shown in FIG. 1 is neutral, the pressure Pz upstream of the meter-in variable restrictors 14a and 14b, that is, in this case, the pump pressure Pd is applied to the first oil chamber 23 of the pressure compensating valve 9a. A second oil chamber 25 provided to the first pressure receiving portion 22 facing the first oil chamber 23 and provided on the side opposed to the first oil chamber 23 is a throttle provided between the inside of the second oil chamber 25 and the port portion 30. Part 41, passage
The tank pressure is communicated with the tank 13 via 17a, 17b, 16a, 16b, the detection ports 15a, 15b, and the tank passages 13a, 13b, and the force applied to the first pressure receiving chamber 22 is equal to the force of the spring 20a and the force of the spring 20a. 2 is greater than the sum of the forces applied to the pressure receiving portions 24,
The main valve element 9a moves downward in FIG. 1, whereby the gap between the pump passage 6a and the pressure supply passages 11a and 11b is cut off. Further, due to the above-described downward movement of the main valve body 9a, another valve body 40 is pressed against the seat surface of the port portion 30 via the spring 20a, and the port portion 30 is closed by the other valve body 40. .
In this neutral state, the metering-in variable throttle portions 14a and 14b are closed, so that the pressure oil supply passage 11 is closed.
Thus, the gap between the load passages 12a and 12b and the load passages 12a and 12b is cut off, and no pressure oil is supplied to the swing motor 4, so that the swing motor 4 is kept stopped.

From such a neutral state, the swing motor 4 is driven independently,
Alternatively, when the flow control valve 8 is switched for the purpose of combined driving of the swing motor 4 and another actuator (not shown), the meter-in variable throttle portion 14a or 14b is opened, and simultaneously with this opening, the load passage 12a or 12b is opened. Detection port
15a or 15b are communicated, the load pressure 16a or 16b of the load passage 12a or 12b, is guided to the port portion 30 via the passage 17a or 17b, the load pressure P _L is of another valve body 40 shown in FIG. 1 Provided in the lower part, whereby another valve element 40 is provided.
1 moves away from the seat surface of the port portion 30, that is, moves upward in FIG. 1, and is pressed against the illustrated upper end portion of another valve body 40, or the main valve body 9a is moved to another position via a spring 20a. In conjunction with the valve body 40, the movable member moves upward in FIG. 1, that is, in the opening direction in which the throttle portion 26 opens. That is, almost simultaneously with the switching operation from the neutral state of the flow control valve 8, the opening between the pump passage 6a and the pressure oil supply passage 11a or 11b and the opening of the variable throttle portion 14a or 14b allow the discharge oil of the hydraulic pump 1 to be reduced. The swing motor 4 is supplied to the swing motor 4 via the pump passage 6a, the pressure oil supply passage 11a or 11b, and the load pressure supply passage 12a or 12b, and the swing motor 4 is driven.

While the swing motor 4 is being driven, the static stability is as described above.
Regarding dynamic stability, unnecessary movement of the main valve element 9a due to the inertia of the main valve element 9a, the force of the spring 20a, or the influence of the fluid flow force, etc., is caused by the second valve element located inside the other valve element 40. The pressure in the oil chamber 25 is regulated by restricting its inflow and outflow through the throttle section 41.

In the first embodiment, as described above, the main valve body 9a is connected to the pump passage 6a and the pressure oil supply passage 11a via the movement of another valve body 40 in accordance with the switching operation of the flow control valve 8 from neutral. 1
1b, the hydraulic oil is immediately supplied to the swing motor 4 to ensure excellent responsiveness with a small time delay in the drive of the swing motor 4 for switching from the neutral state to the flow control valve 8. In addition, when the swing motor 4 is being driven, unnecessary movement of the main valve element 9a due to the influence of the inertia of the main valve element 9a, flow force, or the like is restricted, and dynamic stability is improved. And the rotation motor 4
Operability and the efficiency of work performed via the turning motor 4 can be improved.

FIG. 3 is a longitudinal sectional view showing a configuration of a main part of a directional control valve having a second embodiment of the present invention.

In the second embodiment shown in FIG. 3, a third pressure receiving portion 50 different from the first pressure receiving portion 22 is formed on one end side of the main valve body 9a so as to face the third oil chamber 51. There is a second on the other end of the main valve body 9a.
A fourth pressure receiving portion 52 different from the pressure receiving portion 24 is formed so as to face the fourth oil chamber 53, and the fourth oil chamber 53 is provided with a hydraulic pressure source.
The oil pressure Pi of the oil pressure source 54 can be supplied to the fourth oil chamber 53, and the third oil chamber 52 is connected to the tank 13. Other basic configurations are almost the same as those of the first embodiment shown in FIG. Note that the pressure receiving area of the first pressure receiving section 22 and the pressure receiving area of the second pressure receiving section 24 are set to, for example, the same A, and the pressure receiving area of the third pressure receiving section 50 and the pressure receiving area of the fourth pressure receiving section 52 are set. The pressure receiving area is set to, for example, the same a.

Acting on the main valve body 9a of the pressure compensating valve 9 in the second embodiment thus constructed balancing, when the tank pressure T, the force of the spring 20a to f, Pz · A + T · a = P L · A + Pi A + f (2) Assuming that T = 0, Pz−P _L = (a / A) · Pi + f / A = ΔP ′ (3) Since a, A, and f are constants in the equation (3), the differential pressure ΔP ′ can be selected by appropriately selecting the hydraulic pressure Pi of the hydraulic power source 54. By providing the spring 20a, the shape and dimensions of the spring 20a
That is, even when the force cannot be increased sufficiently, a large differential pressure ΔP 'can be set regardless of the force of the spring 20a. high flow rate can be supplied, which is suitable, in particular hydraulic Shiyoberu such a large load pressure P _L occurs. Other functions and effects are the same as those of the first embodiment.

The pressure of the third oil chamber 51 and the fourth oil pressure 53
May be controlled via an electromagnetic proportional valve or the like. In the case of such a configuration, the pressure difference ΔP ′ can be set more freely, and the flow rate can be easily controlled. Can be.

In the first and second embodiments, the flow restricting means, for example, the throttle valve 41 is provided only on the second oil chamber 25 side, but the movement of the main valve body 9a in the opening direction is not so much restricted. The first oil chamber 23 may be provided with a flow restricting means having a restricting amount of the above. In this configuration, the operating speed of the main valve element 9a in the opening direction can be controlled, and the damping in the opening direction can be performed. Can be increased as long as the above-described responsiveness is not impaired.

〔The invention's effect〕

Since the pressure compensating valve of the present invention is configured as described above, it is possible to improve the responsiveness of the actuator with respect to switching from the neutral state of the flow control valve and to improve the dynamic stability when the actuator is driven. As a result, it is possible to improve the operability of the actuator and the efficiency of work performed via the actuator as compared with the related art.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional view showing the configuration of a directional switching valve having a first embodiment of a pressure compensating valve according to the present invention, and FIG. 2 is a load provided with a directional switching valve having the pressure compensating valve shown in FIG. FIG. 3 is a circuit diagram showing a hydraulic drive device of a civil engineering / construction machine having a sensing system, FIG. 3 is a longitudinal sectional view showing a configuration of a directional control valve showing a second embodiment of the present invention, and FIG. FIG. 5 is a longitudinal sectional view showing the configuration of a directional control valve having a compensating valve,
FIG. 4 is a circuit diagram showing a hydraulic drive device of a civil engineering / construction machine having a load sensing system provided with a directional control valve having a pressure compensating valve shown in FIG. 4, and FIG. FIG. 4 is a characteristic diagram showing a relationship between a spool stroke of a provided flow control valve and an opening area of each part. 5: Directional switching valve, 6: Block body, 6a: Pump passage, 7: Spool, 8: Flow control valve, 9: Pressure compensating valve, 9a: Main valve body, 11a, 11b ... Pressure oil supply passage, 12
a, 12b ... load passage, 13 ... tank, 14a, 14b ... variable throttle part, 20a ... spring, 22 ... first pressure receiving part, 23 ... first oil chamber, 24 ... second Pressure receiving part, 25 ... second oil chamber, 26
…… Throttle part, 30 …… Port part, 40 …… Another valve body, 40a…
... passage, 41 ... throttle part, 50 ... third pressure receiving part, 51 ... third oil chamber, 52 ... fourth pressure receiving part, 53 ... fourth oil chamber, 54
...... Hydraulic source.

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int.Cl. ⁶ , DB name) F15B 11/00-11/22

Claims (4)

(57) [Claims] A main valve body slidably provided in a block body, wherein a pressure oil supply passage provided in the block body and a load passage connected to an actuator are cut or disconnected, or according to an operation amount. A pressure compensating valve which communicates with the opening amount and controls a differential pressure between an upstream pressure of a meter-in side variable throttle portion connecting the pressure oil supply passage and the load passage and a downstream load pressure of the meter-in variable throttle portion. A first pressure receiving portion provided at one end of the main valve body and facing the first oil chamber into which the upstream pressure is introduced; and a first pressure receiving portion provided at the other end of the main valve body,
A second oil chamber facing the second oil chamber into which the downstream load pressure is introduced.
A pressure compensating valve provided to the first pressure receiving portion and the second pressure receiving portion such that the upstream pressure and the downstream load pressure are opposed to each other. A seat surface is formed in a port portion of at least one of the oil chambers of the chamber and the second oil chamber, and another valve body that can be joined to the seat surface and is interlocked with the main valve body; A pressure compensating valve provided with a flow restricting means capable of restricting a flow rate of the pressure oil flowing between the oil chamber and the port. 2. The pressure compensating valve according to claim 1, wherein another valve body is provided so as to be slidable with respect to the main valve body. 3. A spring provided between the main valve body and another valve body for urging the another valve body against a seat surface of the port portion. ). 4. A main valve body having a third pressure receiving portion at one end side different from the first pressure receiving portion to which pressure upstream of the meter-in variable throttle portion is applied, and a meter-in variable throttle at the other end side. A fourth pressure receiving portion different from the second pressure receiving portion to which a downstream side load pressure is applied is provided.
The pressure compensating valve according to any one of (1) to (3).