JPH048902A - pressure compensation valve - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to pressure upstream and downstream load pressure of a flow control valve that constitutes a directional control valve provided in a hydraulic machine equipped with a load sensing system such as a hydraulic excavator. This invention relates to a pressure compensation valve that controls the differential pressure between the

[Conventional technology]

FIG. 4 is a vertical cross-sectional view showing the configuration of a directional control valve having a conventional pressure compensation valve of this type, and FIG. 5 is a load sensing system equipped with a directional control valve having a pressure compensation valve shown in FIG. 4. FIG. 2 is a circuit diagram showing a hydraulic drive device for civil engineering/construction machinery equipped with the following.

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 rate control device 2 that controls the displacement of the variable displacement hydraulic pump 1, and a pressure oil discharged from the pump 1. A relief valve 3 that regulates the pressure of It is equipped with a directional switching valve 5 that controls the flow of oil.

As shown in FIG. 4, the above-mentioned directional switching valve 5 includes a block body 6 forming a main body, a flow rate control valve 8 having a sub-L7 that slides inside the block body 6, and a flow rate control valve 8 having A first pressure receiving part 22 is provided on the upstream side of the meter-ine variable throttle part 14, 1.14b in the flow rate control valve 8, and a second pressure receiving part 22 is provided with a pressure Pz upstream of the meter-ine variable throttle part 14, 1.14b, and a second pressure receiving part 22 is provided with a downstream load pressure PL. It has a pressure receiving part 24, and the pressure P upstream of these parts is
z and the downstream load pressure PL, that is, the front and rear differential pressure Pz
PL is controlled by the spring 20 to the set differential pressure ΔP', and fluctuations in pump pressure Pd and load pressure P during combined operation are controlled.
In order to supply a substantially constant flow rate to the swing motor 4 regardless of fluctuations in L, a pressure compensation valve 9 having a main valve body 9a slidably provided on the block body 6 and a pressure compensation valve 9 provided downstream of the flow rate control valve 8 are provided. A shuttle valve 10 is provided.

The above-mentioned block body 6 has two pressure oil supply passages n11a and llb that are connected to the pump 1, and a load passage that can communicate with these pressure oil supply passages 11a and llb, respectively, and is connected to the swing motor 4. 12a, 12b, and a tank passage 13a that can communicate with these load passages 12a, 12b.
13b is provided. In addition, the above-mentioned spool 7 is connected with the pressure oil supply passage 11a and the load passage 12a, or the pressure oil supply passage 11b and the load passage 12b, each of which is opened by an amount corresponding to the stroke of the spool 7. variable throttle parts 14a, 14b of the meter tine, detection ports 15a, 15b, which are provided downstream of these variable throttle parts 14a, 14b, to detect the load pressure PL of the swing motor 4 shown in FIG. port 15a,
15b, passages 17a and 17b that communicate with these passages 16a and 16b, and a passage 18 that can communicate with these passages 17a and 17b. The above-mentioned passages 16a, 16b, 17a, 17b
, 18 are connected to a passage 19 connected to the shuttle valve 10 shown in FIG. 5, and the pump flow rate control device 2 shown in FIG. It constitutes a transmission path for transmitting information to one drive section. Note that the pump pressure Pd is guided to the other driving part of the pump flow rate control device 2, and this bob flow rate control device 2 is controlled so that the differential pressure between the pump pressure Pd and the load pressure PL is a preset pressure, i.e. The displacement of the pump 1 is controlled so as to balance the pressure ΔP corresponding to the force of the spring 2a. That is, pump pressure Pd - load pressure P
The flow rate of the pump 1 is controlled so as to always maintain L+ΔP.

When the spool 7 is in an operating state near the neutral position, the detection ports 15a and 15b are connected to the tank passage 13a.
, 13b, and the aforementioned passages 16a, 16b, 1.7
A, 17b, and 18 are connected to the tank passage 13a.
, 13b.

In the directional switching valve 5 configured as described above, the switching timing of the variable throttle portions 14a, 14b, the detection ports 15a, 15b, etc. relative to the spool stroke of the flow rate control valve 8 is adjusted so that the spool 7 is controlled to drive the swing motor 4 independently. Assuming that it is moved from neutral to the right in Figure 4,
This will be explained with reference to a special drawing showing the relationship between the stroke of the spool 7 and the opening area shown in the figure. In addition, in FIG. 6, a characteristic line 20a indicates the opening area between the detection port 15a and the tank passage 13a, and a characteristic line 20b indicates the opening area between the passage 18 and the passage 17b.
That is, the characteristic line 20c shows the opening area between the tank passage 13b, the characteristic line 20c shows the opening area between the detection port 15a and the load passage 12a, and the characteristic line 20d shows the opening area of the variable throttle section 14a.

When the spool 7 of the flow control valve 8 moves to the right from the state shown in FIG. 4, the detection port 15a and the tank passage 13a are first cut off, as shown by the characteristic line 20a in FIG. At this time, as shown by the characteristic line 20b,
18 communicates with the tank passage 13b via passages 17b, 16b and the detection port 15b. Further, when the spool 7 moves to the right, the passage 18 and the passage Nl 7b are cut off, and the discharge passage that had been formed so far to connect the transmission passage including the passage 18 and the tank passage 13b. disappears. Furthermore, when the spool 7 moves to the right, the detection port 15a opens to the load passage 12a as shown by the characteristic line 20c, and the load pressure PL of the swing motor 4 shown in FIG. , 17a
, 18, the shuttle valve 10, and is transmitted to one drive section of the pump flow rate control device 2 via the passage N19 shown in FIG. As a result, the discharge pressure Pd of the pump 1 is Pd=PL+
The pressure is increased to ΔP. When the spool 7 further moves from the above-mentioned state, the variable throttle part 14a of the meterine opens as shown by the characteristic line 20d in FIG.
Pressure oil supplied from the hydraulic pump 1 shown in the figure via the pressure compensation valve 9 is guided to the swing motor 4 via the pressure oil supply passage 11a, the variable throttle part 14a, and the load passage #12a shown in FIG. , this swing motor 4 is activated and can drive a rotating structure (not shown). At this time, in the variable throttle part 14a, ΔP is calculated with respect to the opening amount of the variable throttle part 14.[Currently, since the pressure compensation valve 9 is independently driven by the swing motor 4, pressure compensation is not performed, and ΔP=I (Pz= Pd)
-PL). ), where C is a constant and A is the opening area of the variable restrictor 14a, the flow rate Q supplied to the swing motor 4 is Q or c-AfK7 (1). In addition, the pressure compensation valve 9 does not function when the rotary motor 4, which is the swing motor, is driven alone, but when the swing motor 4 and other actuators (not shown) are driven in combination, the pressure compensation valve 9 is variable in order to realize a good combined drive of these actuators. Control is performed so that the differential pressure Pz-Pt across the throttle portions 14a and 14b remains constant at the differential pressure ΔP' set by the spring 20. In this case, when the pump pressure Pd increases with the combined operation,
The flow rate passing through the variable throttle parts 14a, 14b of the meter tine increases, and the upstream pressure Pz of the variable throttle parts 14a, 14b is introduced into the first oil chamber 23 via the passage 26, and is added to the first pressure receiving part 22. be done. In addition, the variable aperture parts 14a and 14b
The port pressure, that is, the load pressure Pt, which is the downstream pressure of , is introduced into the second oil chamber 25 from the port portion 30 via the passages B15a, 16a, and 17a, and is applied to the second pressure receiving portion 24. Here, the main valve body 9 acting on the main valve body 9a
Due to the downward movement of a in FIG.
The flow rate is determined by balancing the flow rate so that it becomes P'.

[Problem to be solved by the invention]

By the way, in the conventional pressure compensating valve 9 described above, statically, as described above, it operates in a balanced manner according to the differential pressure ΔP when the swing motor 4 is driven alone, and according to the differential pressure ΔP' when the swing motor 4 is driven in combination. With regard to dynamic movement, the pressure compensation control tends to become unstable due to the influence of the inertia of the main valve pair 9a, the force of the spring 20, the fluid flow force, etc. therefore,
Considering this dynamic stability, it is necessary to provide the fluid passage with a restrictor or the like that limits the flow rate. That is, the fourth
A constriction part is provided in the passage 27 that leads the pressure Pz upstream of the supply passages 11a and llb shown in the figure to the first oil chamber 23, and it also communicates with the second oil chamber 25 and is a passage with a load pressure Pt. It becomes necessary to increase damping by providing a constriction section in the port section 30. However, such a restricting portion is limited to the opening direction of the main valve body 9a of the pressure compensation valve 9, that is, the upward movement in FIG. 4, and the closing direction of the main valve body 9a, that is, the downward movement in FIG. It acts as a resistance to any of the operations. 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
decreases to the tank pressure, the main valve body 9a resists the force of the spring 20 and is positioned to the lower limit in FIG. 4, and the throttle section 26 becomes fully closed. If only the amount of throttle is assumed in consideration of the
The operation in the opening direction, that is, the opening operation of the throttle section 26, is delayed, and the response of the swing motor 4 to the switching operation of the slanted flow control valve 8 deteriorates, resulting in deterioration in the operability of actuators such as the swing motor, and This will result in a decrease in the efficiency of the work being done.

The present invention has been made in view of the above-mentioned actual situation in the prior art, and its objectives are to improve the responsiveness of the actuator to switching the flow control valve from neutral and to improve the dynamic stability when the actuator is driven. An object of the present invention is to provide a pressure compensating valve that can realize the following.

[Means to solve the problem]

In order to achieve this object, the present invention has a main valve body that is slidably provided in a block body, and the pressure oil supply passage provided in the block body and the load passage connected to the actuator are shut off or operated. A pressure that controls the differential pressure between the upstream pressure of the meter-in side variable throttle section that communicates with the opening amount depending on the amount of pressure oil supply passage and the load passage, and the downstream load pressure of the meter-in side variable throttle section that connects the pressure oil supply passage and the load passage. a compensation valve, which is provided at one end of the main valve body, and includes a first pressure receiving portion facing a 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 pressure receiving part facing the second oil chamber into which the downstream load pressure is introduced, and the upstream pressure and the downstream load pressure are connected to the first pressure receiving part and the second pressure receiving part. in the pressure compensating valve provided so as to oppose each other, a seat surface is formed in a port portion of at least one of the first oil chamber and the second oil chamber, and the seat surface is connectable to the first oil chamber and the second oil chamber. The main valve body is configured to include another valve body that can interlock with the main valve body, and a flow rate limiting means that can limit the flow rate of pressure oil flowing between the at least one oil chamber and the port portion.

[Effect]

Since the present invention is configured as described above, when the flow control valve is in the neutral state, the pressure upstream of the variable restrictor of the meter tine, that is, the pump pressure Pd, is applied to the first pressure receiving pressure of the main valve body facing the first oil chamber. The oil chamber of the force cylinder 2 becomes tank pressure via the flow rate limiting means provided between the inside of this second oil chamber and the port section, and the main valve body controls the pressure between the pump passage and the pressure oil supply passage. At the same time, the variable throttle section of the meterine is closed to cut off the pressure oil supply passage and the load passage connected to the actuator, and in conjunction with the closing operation of the main valve element, another valve element closes the seat of the port section. This port portion is closed by contacting the surface.

When the flow rate control valve is switched from such a neutral state, the load pressure PL is applied to another valve element simultaneously with the opening of the variable throttle part of the meterine, and this other valve element moves away from the seat surface of the port part. However, the main valve body moves in the opening direction in conjunction with this other valve body.

That is, the pump passage and the pressure oil supply passage open almost simultaneously with the neutral switching operation of the flow control valve, and the pressure oil supply passage and the load passage communicate with each other through the pump passage, the pressure oil supply passage, and the load passage. Pressure oil can be supplied to the actuator to drive the actuator, and the responsiveness of the actuator to switching from the neutral state of the flow control valve can be improved.

The static stability of such an actuator during operation is determined by the force exerted by the pressure upstream of the variable throttle section of the meterine applied to the first pressure receiving section of the main valve body, and the force exerted on the second pressure receiving section of the main valve body. The difference between the force applied to the pressure receiving part and the pressure downstream of the variable throttle part of the meterine is maintained at a predetermined value in the same manner as in the prior art, and dynamic stability is maintained by the movement of the main valve body. Since the flow of pressure oil between the inside of the second oil chamber and the port is restricted by the flow rate restriction means, unnecessary movement of the main valve body is restricted, and its stability can be improved. .

[Example] Hereinafter, an example of the pressure compensating valve of the present invention will be described based on the drawings.

FIG. 1 is a longitudinal cross-sectional view showing the configuration of a directional control valve having a first embodiment of the present invention, and FIG. 2 shows a load sensing system equipped with a directional control valve having a pressure compensation valve shown in FIG. FIG. 2 is a circuit diagram showing a hydraulic drive system for civil engineering and construction machinery.

The directional control valve 5 shown in FIG. 1 is drawn in correspondence with the directional control valve shown in FIG. A flow rate control valve 8 having a slidably provided spool 7, a pump passage 6a, pressure oil supply passages 11a and 11b that can communicate with the pump passage 6a, and a meter-in variable throttle part 14a.
, 14b, and tank passages 13a, 13b, detection ports 15a, 15b, passages 16a, 16b, 17a, 17b, shuttle valve 10, and a passage 18.

The hydraulic circuit shown in FIG. 2 is drawn in correspondence with the hydraulic circuit shown in FIG. A pump flow rate control device 2 that controls the displacement of the variable displacement hydraulic pump l so that the flow rate corresponds to the differential pressure ΔP, and a relief valve 3. A passage 19 is provided that guides the swing motor 4 exemplified as an actuator, a tank 13, and a load pressure P to the pump flow rate control device 2.

Further, the pressure compensating valve 9 of this embodiment is similar to the one shown in FIG.
a passage 27 that guides the pressure upstream of 4b to the first oil chamber 23;
The first oil chamber facing the first oil chamber 23 and receiving the upstream pressure Pz
facing the pressure receiving part 22 and the second oil chamber 25, the load pressure PL
The main valve element 9a is provided with a second pressure receiving part 24 that receives the pressure.

The second pressure receiving part 24 of the pressure compensating valve 9 is formed in a cylindrical shape, and another valve element 4 is disposed inside this cylindrical part.
0 is slidably provided. A seat surface is formed in the port portion 30, and the above-mentioned another valve body 40 can be joined to this seat surface. This other valve body 40 is provided with a flow rate restriction means capable of restricting the flow rate of the pressure oil flowing between the second oil chamber 25 and the port portion 30, for example, a throttle portion 41, and the inside of the other valve body 40 is A passage 40a is provided which communicates a portion of the second oil chamber 25 located at the outer side of the valve body 40 with a portion of the second oil chamber 25 located outside of another valve body 40. Further, a spring 20a is provided between the inner part of the other valve element 40 and the second pressure receiving part 24 of the main valve element 9a, which biases the other valve element 40 against the seat surface of the port part 30. The main valve body 9a and another valve body 40 can communicate with each other via this spring 20a.

In the embodiment configured in this way, the static movement is equivalent to the prior art shown in FIG. 4.5.6, and when the swing motor 4 is driven independently,
Due to the failure of the flow control valve 8, the differential pressure Pz across the flow control valve 8
The displacement of the variable displacement hydraulic pump 1 is controlled so that the force due to -PL, that is, PdPt (=ΔP) and the force of the spring 2a of the pump flow rate control device 2 are balanced.
1) A flow rate corresponding to the differential pressure ΔP shown by the formula is supplied to the swing motor 4, and the swing motor 4 is driven. Also, when the swing motor 4 and an actuator (not shown) are driven in combination, the pressure compensation valve 9 functions. The differential pressure P across the flow control valve 8
The force due to z-PL=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, and the combination of this swing motor 4 and other actuators is controlled. drive can be realized.

When the flow rate control valve 8 shown in FIG.
That is, in this case, the pump pressure Pd is applied to the first pressure receiving part 22 facing the first oil chamber 23 of the main valve body 9a of the pressure compensating valve 9, and is provided on the side opposite to the first oil chamber 23. The second oil chamber 25 includes a throttle section 41 provided between the inside of the second oil chamber 25 and the port section 30, passages 17a and 17b,
16a, 16b, detection ports 15a, 15b, and tank passages 13a, 13b, it communicates with the tank 13 and becomes tank pressure, and the force applied to the first pressure receiving part 22, the force of the spring 20a, and the second pressure receiving part 24 becomes larger than the sum of the forces applied to the main valve body 9a, and moves downward in FIG.
llb 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 this other valve body 40 is pressed against the seat surface of the port portion 30.
0 closes the port portion 30.

In this neutral state, the variable throttle parts 14a and 14b of the meterine are closed, thereby cutting off the pressure oil supply passages 11a and llb and the load passages 12a and 12b, and the swing motor 4 is supplied with pressure oil. Not supplied and kept in a stopped state.

When the flow rate control valve 8 is switched from such a neutral state with the intention of driving the swing motor 4 alone or in combination with the swing motor 4 and another actuator (not shown), the variable throttle section 14a or 1.4 of the meterine is switched. b opens, but at the same time as this opening, the load passage 12a or 12b
The detection port 15a or 15b communicates with the load pressure 16a or 16 of the load passage 12a or 12b.
b, port section 30 via passage 17a or 17b
The load pressure PL is guided to another valve body 4 shown in FIG.
0, thereby causing another valve element 40 to move away from the seating surface of the port portion 30, that is, upwardly in FIG. Alternatively, the main valve body 9a moves upward in FIG. 1 in conjunction with the other valve body 40 via the spring 20a, that is, in the opening direction in which the throttle portion 26 opens. In other words, the flow control valve 8
Almost simultaneously with the switching operation from neutral, the gap between the pump passage 6a and the pressure oil supply passage 11a or llb is opened, and the variable throttle section 14a or 14b is opened.
The oil discharged from the hydraulic pump 1 is supplied to the swing motor 4 via the pump passage 6a, the pressure oil supply passage 11a or llb, and the load pressure supply passage 12a or 12b, and the swing motor 4 is driven.

While the swing motor 4 is being driven,
Static stability is as described above, but regarding dynamic stability, unnecessary main valve body 9a due to the influence of inertia of main valve body 9a, force of spring 20a, flow force of fluid, etc. The movement of the pressure oil in the second oil chamber 25 located inside another valve body 40 is regulated by restricting the inflow and outflow of the pressure oil through the throttle portion 41 .

In this first embodiment, as described above, when the flow rate control valve 8 is switched from neutral, the main valve body 9a is connected to the pump passage 6a and the pressure oil supply passage 11a or llb, so that pressure oil is immediately supplied to the swing motor 4, and excellent responsiveness with little time delay in driving the swing motor 4 in response to switching of the flow control valve 8 from neutral is ensured. In addition, while the swing monitor 4 is being driven, unnecessary movement of the main valve body 9a due to the influence of inertia of the main valve body 9a, flow force, etc. is restricted, and dynamic stability is improved. This makes it possible to improve the operability of the swing motor 4 and improve the efficiency of work performed via the swing motor 4.

FIG. 3 is a longitudinal sectional view showing the configuration of the main parts 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 section 50 different from the first pressure receiving section 22 is provided at one end side of the main valve body 9a.
A fourth pressure receiving part 52, which is different from the second pressure receiving part 24, is formed on the other end side of the main valve body 9a, facing the fourth oil chamber 53. At the same time, a hydraulic source 54 is connected to the fourth oil chamber 53 so that the hydraulic pressure P1 of this hydraulic source 54 can be supplied to the fourth oil chamber 53, and the third oil chamber 51 is connected to the tank 13. It is connected to. The other basic configuration is almost the same as the first embodiment shown in FIG. 1 described above. Note that the pressure receiving area of the first pressure receiving part 22 and the pressure receiving area of the second pressure receiving part 24 are set to the same value A, for example, and the pressure receiving area of the third pressure receiving part 50 and the pressure receiving area of the fourth pressure receiving part 52 are set to be the same. The pressure receiving area is also set to the same value a, for example.

Pressure compensation valve 9 in the second embodiment configured in this way
The balance of forces acting on the main valve body 9a is the tank pressure T
, if the force of the spring 20a is f, then Pz-A+T-a=P
Considering L-A+Pi-a+fT=O, Pz-PL = (a/A) -P f +f/A=Δ
P' (3). This (3
), since a, A, and f are constants, the differential pressure ΔP' can be selected by appropriately setting the hydraulic pressure Pi of the hydraulic source 54.
By providing the spring 20a, even if the shape and dimensions of the spring 20a, that is, the force cannot be made sufficiently large, a large differential pressure ΔP' can be set regardless of the force of the spring 20a, and this allows the flow rate control valve to 8 can supply a sufficiently large flow rate to the swing motor 4, and is particularly suitable for hydraulic excavators and the like in which a large load pressure PL is generated. Other effects are the same as those of the first embodiment described above.

Note that the pressure in the third oil chamber 51 and the fourth oil chamber 5 described above
The magnitude of the pressure in step 3 may be controlled via an electromagnetic proportional valve, etc. With this configuration, the differential pressure ΔP' can be set more freely, and the flow rate can be easily controlled. be able to.

In the first and second embodiments described above, the flow rate restricting means, for example, the throttle portion 41 is provided only on the second oil chamber 25 side;
A flow rate limiting means having an amount of restriction that does not restrict the movement of the main valve body 9a in the opening direction may also be provided on the first oil chamber 23 side.
The operating speed in the opening direction can also be controlled, and the damping in the opening direction can also be increased within a range that does not impair the above-mentioned responsiveness.

〔Effect of the invention〕

Since the pressure compensating valve of the present invention is configured as described above,
It is possible to improve the responsiveness of the actuator to switching from the neutral state of the flow control valve and to improve the dynamic stability when driving the actuator. It is possible to improve the efficiency of work performed through

[Brief explanation of the drawing]

FIG. 1 is a longitudinal sectional view showing the structure of a directional control valve having a first embodiment of the pressure compensating valve of the present invention, and FIG. 2 is a load bearing equipped with a directional control valve having the pressure compensating valve shown in FIG. A circuit diagram showing a hydraulic drive system for civil engineering and construction machinery equipped with a sensing system, FIG. 3 is a vertical sectional view showing the configuration of a directional control valve showing a second embodiment of the present invention, and FIG. 4 is a conventional pressure 5th longitudinal cross-sectional view showing the configuration of a directional control valve having a compensation valve;
The figure is a circuit diagram showing a hydraulic drive system for civil engineering and construction machinery equipped with a load sensing system equipped with a directional control valve having the pressure compensation valve shown in Fig. 4, and Fig. 6 is a circuit diagram showing the directional control valve shown in Fig. 1. It is a characteristic diagram which shows the relationship between the spool stroke of the flow control valve provided, and the opening area of each part. 5...Directional switching valve, 6...Block body, 6a...Pump passage, 7...Spool, 8...Flow rate control valve , 9...Pressure compensation valve, 9a...Main valve body, lla, 11b...
...pressure oil supply passage, 12a, 12b...load passage, 13...tank, 14a, 14b,
-..., variable aperture section, 20a... spring, 22
・Old...first pressure receiving part, 23...first oil chamber,
24...Second pressure receiving part, 25...Second pressure receiving part
oil chamber, 26... throttle section, 3o... port section, 40... another valve body, 40a old... passage, 41... Throttle part, 50...Third pressure receiving part, 51...Third oil chamber, 52...
・Fourth pressure receiving part, 53...Fourth oil chamber, 54・
...Hydraulic power source. Figure 2? P a jlI5 Figure 6 Kufur/? straw 7

Claims (4)

[Claims] (1) It has a main valve body that is slidably installed in the block body, and the pressure oil supply passage provided in the block body and the load passage connected to the actuator are shut off or communicated by an opening amount depending on the amount of operation. The pressure compensating valve controls the differential pressure between the upstream pressure of the meter-in side variable throttle section connecting the pressure oil supply passage and the load passage and the downstream load pressure of the meter-in side variable throttle section, which A first pressure receiving part is provided at one end of the valve body and faces the first oil chamber into which the above-mentioned upstream pressure is introduced, and a first pressure receiving part is provided at the other end of the main valve body and into which the above-mentioned downstream load pressure is introduced. and a second pressure receiving part facing the second oil chamber, and the first pressure receiving part and the second pressure receiving part are provided with:
In the pressure compensation valve in which the upstream pressure and the downstream load pressure are applied so as to oppose each other, a seat surface is provided at a port portion of at least one of the first oil chamber and the second oil chamber. another valve body that can be formed and joined to the seat surface and interlocked with the main valve body, and a flow rate restriction means that can restrict the flow rate of pressure oil flowing between the inside of the at least one oil chamber and the port section. A pressure compensation valve characterized by being provided with. (2) The pressure compensating valve according to claim (1), characterized in that another valve body is provided slidably with respect to the main valve body. (3) Claim (1) or (2) characterized in that a spring is provided between the main valve element and another valve element to bias the other valve element against the seat surface of the port portion. pressure compensation valve. (4) The main valve body has a third pressure receiving part on one end side which is different from the first pressure receiving part to which the upstream pressure of the variable throttle part of the meter-ine is applied, and the other end side has a third pressure receiving part to which the upstream pressure of the variable throttle part of the meter-ine is applied. A fourth pressure receiving part different from the second pressure receiving part to which the downstream load pressure is applied.
Claims (1) to (3) characterized by having a pressure receiving part of
) Pressure compensating valve described in any of the above.