JP3144914B2 - Hydraulic control valve device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control valve device used for a hydraulic drive device of a construction machine, and more particularly, to a spool type directional control valve which has a pressure compensation function and a load check function and has a small pressure loss. The present invention relates to a hydraulic control valve device.

[0002]

2. Description of the Related Art In construction machines such as hydraulic excavators, a hydraulic drive device for driving a plurality of actuators with hydraulic oil discharged from one hydraulic pump is generally used. Japanese Patent Application Laid-Open No. 11706 discloses a load sensing type hydraulic drive device. In this hydraulic drive device, the discharge pressure oil from the hydraulic pump is supplied to a plurality of actuators via a plurality of spool type direction switching valves, and a variable displacement type hydraulic pump is used as the hydraulic pump. The flow rate is controlled according to the maximum load pressure of the plurality of actuators. Further, a pressure compensating valve for controlling the pressure difference between the front and rear of the variable throttle portion of the direction switching valve is arranged upstream of the plurality of direction switching valves so that the operation speed of the plurality of actuators does not change due to a load change. . Further, a load check valve for preventing backflow of the pressure oil is disposed between each pressure compensating valve and each direction switching valve.

In a load sensing type hydraulic drive device, during a combined operation of simultaneously driving a plurality of actuators, the discharge flow rate of the hydraulic pump is insufficient, the discharge pressure of the hydraulic pump is reduced, and the pressure compensation valve does not operate. Sometimes. In this case, a large amount of pressure oil is supplied to the light-load-side actuator, and the heavy-load-side actuator is not driven, so that an appropriate combined operation cannot be performed. In order to solve this problem, in the hydraulic drive device described in Japanese Patent Application Laid-Open No. Sho 60-11706, instead of a spring normally provided as means for indicating a target compensation differential pressure of a pressure compensating valve, a hydraulic pump is used. The differential pressure between the discharge pressure and the maximum load pressure is applied, so that when the pump discharge flow rate is insufficient, the target compensation differential pressure is reduced in response to the decrease in the differential pressure. Operation is ensured.

On the other hand, when a hydraulic drive is constructed by combining a spool-type directional control valve, a pressure compensating valve, and a load check valve as described above, these three components are used for the purpose of reducing the number of pipes and making them compact. It has been considered that one valve device is constructed by incorporating one valve into one block.
There is a valve device described in Japanese Patent No. 4402.

That is, in the valve device described in Japanese Patent Application Laid-Open No. 02-134402, a horizontal hole is formed through the valve block, and a vertical hole is formed perpendicularly to the center of the horizontal hole. Insert the spool of the directional control valve slidably,
The balance piston of the pressure compensation valve is slidably housed in the vertical hole. A pump port, a feeder passage, and an actuator port are formed in the valve block as part of a main circuit that supplies pressure oil from a hydraulic pump to the actuator, and the spool of the directional control valve is connected to the feeder passage and the actuator port. A possible main variable throttle is formed, and an inner hole communicating with the pump port and the feeder passage is formed in the balance piston, and the inner hole and the feeder passage are connected via a variable throttle for pressure compensation. Also,
The valve body of the load check valve is slidably housed in the inner hole of the balance piston.

Japanese Patent Application Laid-Open No. 58-501781 discloses a seat valve type hydraulic control valve device, which is composed of a combination of a seat valve and a pilot control valve.

[0007]

Conventionally, as a hydraulic control valve device for construction machinery, a device provided with a spool-type directional control valve has been used for many years. Is reliable and easy to design. However, the control valve device having the conventional spool-type directional control valve has the following problems.

As described above, Japanese Patent Application Laid-Open No. Sho 60-11706
In the hydraulic control valve device described in Japanese Patent Application Laid-Open No. 2002-134402 and Japanese Patent Application Laid-Open No. 02-134402, three valves, that is, a direction switching valve, a pressure compensating valve, and a load check valve are arranged in a main circuit. Therefore, the pressure oil supplied to the actuator from the hydraulic pump receives the resistance each time it passes through these three valves, causing a pressure loss, resulting in a large energy loss and uneconomical.

Further, in the valve device described in Japanese Patent Application Laid-Open No. 02-134402, it is necessary to form a number of complicated pressure receiving chambers, passages and the like in the balance piston of the pressure compensating valve. That is, it is necessary to form pressure receiving chambers at both ends of the balance piston independently of the pump port to introduce the inlet pressure and the outlet pressure of the main variable throttle, and to make the target compensation differential pressure of the pressure compensating valve variable. It is necessary to additionally install two pressure receiving chambers. Further, it is necessary to form an inner hole for accommodating the load check valve element of the main circuit inside the balance piston. For this reason, compared with a valve device having only a load check valve without a pressure compensation function, the circumference of the balance piston and the balance piston itself become large, the valve block becomes longer in the axial direction of the balance piston, and the outer shape of the valve block becomes larger. Become. Also, the production of the valve block becomes complicated.

On the other hand, the hydraulic control valve device disclosed in Japanese Patent Application Laid-Open No. 58-501781 uses a seat valve type instead of a spool type directional switching valve, and cannot take advantage of the above-mentioned advantages of the spool type. Further, in this hydraulic control valve device, use of a seat valve in combination with a spool-type directional switching valve is not considered at all.

A first object of the present invention is to provide a hydraulic control valve device provided with a pressure compensation function and a load check function and having a small pressure loss in a hydraulic control valve device provided with a spool type directional switching valve. It is.

A second object of the present invention is to provide a hydraulic control valve device provided with a pressure compensation function and a load check function, which is compact and easy to manufacture, in a hydraulic control valve device provided with a spool type directional control valve. It is to be.

[0013]

In order to achieve the first and second objects, according to the present invention, a housing, a pump port formed in the housing and connected to a hydraulic pressure source, and the housing are provided. At least one spool valve body slidably disposed in the housing, a load passage formed in the housing, the feeder passage communicating with the pump port, and a load port connected to an actuator; Located between the load passage,
A hydraulic control valve device having a main variable throttle that changes an opening area in accordance with an amount of movement of the spool valve element, wherein the main port is connected to the main variable throttle via the feeder passage in accordance with a pressure difference across the main variable throttle. Auxiliary flow rate control means for auxiliaryly controlling the flow rate of the pressure oil supplied to the variable throttle,
The auxiliary flow rate control means includes: (a) a seat valve body slidably disposed in the housing between the pump port and the feeder passage; and a seat valve body, the movement of the seat valve body. A control variable throttle that changes an opening area in accordance with an amount of the valve; and a pilot passage that connects the pump port to the feeder passage through the control variable throttle. The seat valve body is controlled by a pilot flow rate that passes through the pilot passage. A seat valve for determining the amount of movement of the seat;
(B) a pilot control means disposed in the pilot passage, the pilot control means controlling the pilot flow rate in accordance with the differential pressure across the main variable throttle.

Preferably, the hydraulic control valve device further includes a fixed block for holding the seat valve body in the housing via a spring, and the pilot control means is a pilot control valve incorporated in the fixed block. is there. In this case, it is preferable that the pilot control valve includes a pilot valve disposed in parallel with the spool valve.

Preferably, the hydraulic control valve device further includes a check valve disposed in the pilot passage for preventing a backflow of the pressure oil in the pilot passage.

In the above-mentioned hydraulic control valve device, preferably, the pilot control means closes a slidable pilot valve body and an urging force corresponding to the differential pressure across the main variable throttle to the pilot valve body. Biasing means for applying in the valve direction, and target compensation differential pressure indicating means for applying a predetermined urging force to the pilot valve body in the valve opening direction and indicating a target value of the differential pressure across the main variable throttle. .

In this case, preferably, the target compensation differential pressure indicating means includes a spring acting on the pilot valve body,
The preset force of the spring is applied in the valve opening direction as the predetermined urging force. Further, the target compensation differential pressure indicating means includes:
Preferably, the apparatus further includes operating means for adjusting a preset force of the spring.

Preferably, the target compensation differential pressure indicating means includes a hydraulic pressure generating means for generating a predetermined hydraulic pressure as the predetermined urging force. The hydraulic pressure generating means preferably includes a pressure receiving chamber into which a variable pressure is introduced.

According to the present invention, in order to achieve the first and second objects, a housing, a pump port formed in the housing and connected to a hydraulic pressure source, A plurality of direction switching valve devices each controlling the flow rate of the pressure oil supplied to the actuator, wherein the plurality of direction switching valve devices are each slidably disposed in the housing and a spool valve body. A load passage formed in the housing and having a feeder passage communicating with the pump port and a load port connected to the actuator; and a movement of the spool valve body located between the feeder passage and the load passage. A hydraulic control valve device having a main variable throttle that changes an opening area according to an amount of the main variable throttle. Auxiliary flow rate control means for auxiliary control of the flow rate of pressure oil supplied from the pump port to the main variable throttle through the feeder passage in accordance with the pressure difference between the front and rear, and the auxiliary flow rate control means (A) a seat valve body slidably disposed in the housing between the pump port and the feeder passage, formed in the seat valve body, and having an opening area according to the amount of movement of the seat valve body And a pilot passage for connecting the pump port to the feeder passage through the controllable throttle, and a moving amount of the seat valve element is determined by a pilot flow rate passing through the pilot passage. A seat valve; and (b) pilot control means disposed in the pilot passage and controlling the pilot flow rate in accordance with a pressure difference across the main variable throttle. The pilot control means further includes means for detecting a maximum pressure among the pressures of the load ports of the plurality of directional control valve devices as a maximum load pressure, wherein the pilot control means includes a pump port as a target value of the differential pressure across the main variable throttle. And a target compensation differential pressure indicating means for indicating a value corresponding to a differential pressure between the pressure and the maximum load pressure.

In the above-mentioned hydraulic control valve device, preferably, the pilot control means of each of the plurality of direction switching valve devices includes a slidable pilot valve body and a front-rear difference between the pilot valve body and the main variable throttle. First biasing means for applying a biasing force in accordance with the pressure in the valve closing direction, wherein the target compensation differential pressure indicating means includes a pilot valve body having a pressure between the pump port pressure and the maximum load pressure. A second urging means for applying an urging force according to the differential pressure in the valve opening direction is included.

[0021]

In the hydraulic control valve device of the present invention constructed as described above, when the spool valve element is operated from the neutral position, the main variable throttle opens, and most of the pressure oil in the pump port passes through the seat valve as the main flow rate. And a part of the pressure oil in the pump port flows through the pilot passage of the seat valve as the pilot flow rate and flows out to the feeder passage. These main flow and pilot flow merge again in the feeder passage, and the merged pressure oil passes through the main variable throttle and is supplied to the load port. Further, when the pressure oil passes through the main variable throttle, a differential pressure across the main variable throttle is generated, and the pilot control means controls the pilot flow rate according to the differential pressure across the main variable throttle. On the other hand, in the case of the seat valve, similarly to the seat valve described in Japanese Patent Application Laid-Open No. 58-501781, the action of a control variable throttle that changes the opening area according to the amount of movement of the seat valve element causes the seat to vary according to the pilot flow. Determine the amount of movement of the valve. For this reason, when the pilot flow rate is controlled according to the differential pressure across the main variable throttle,
Accordingly, the amount of movement of the seat valve body is adjusted, and the main flow rate passing through the seat valve is adjusted. By adjusting the main flow rate, the flow rate of the pressure oil supplied to the main variable throttle is additionally controlled, and the differential pressure across the main variable throttle is also controlled. That is, the pressure difference before and after the main variable throttle is maintained at a predetermined value regardless of the fluctuation of the load pressure or the supply pressure, and the pressure compensation function is performed.

Further, when the load increases and the load pressure becomes higher than the supply pressure, and the pressure oil tries to flow backward, the pilot flow rate becomes zero, so that the movement amount of the seat valve body becomes zero, and the seat valve is fully discharged. Close. Therefore, the backflow of the pressure oil is prevented, and the load check function is performed. Further, in the hydraulic control valve device of the present invention, only two valves, a seat valve and a main variable throttle, are arranged in the pump port, the feeder passage, and the load passage constituting the main circuit. The pressure loss when passing through is reduced, and the energy loss can be reduced.

Further, in the hydraulic control valve device according to the present invention, the seat valve element of the seat valve is disposed at a position in the housing where the load check valve of the conventional valve device without the pressure compensation function is provided. Separate pilot control means is provided,
The pilot flow rate for determining the moving amount of the seat valve element is controlled by the pilot control means. For this reason, the configuration of the seat valve around the seat valve element is simplified, and the axial length of the seat valve element in the portion where the seat valve element is located in the housing does not become large, and the housing becomes compact, and the housing becomes compact. Is easy to manufacture.

[0024] By configuring the pilot control means with a pilot control valve incorporated in a fixed block holding the seat valve element, the pilot control means can be arranged using the fixed block. By disposing the pilot valve element of the pilot control valve in parallel with the spool valve element, the fixed block can be made compact.

By arranging a check valve for preventing the backflow of the pressure oil in the pilot passage, a load check function with higher liquid tightness can be obtained.

In the hydraulic control valve device according to the present invention, wherein the pilot control means includes the urging means and the target compensation differential pressure indicating means, when the load pressure or the supply pressure fluctuates, the pilot valve body is provided by the urging means. The pilot flow rate is controlled by controlling the balance between the urging force according to the differential pressure across the variable throttle and the urging force provided by the target compensation differential pressure indicating means, and the seat valve is controlled by the target compensation differential pressure indicating means. The differential pressure across the main variable throttle is controlled to match the differential pressure.

The simplest form of the target compensation differential pressure indicating means is a spring. In this case, the preset force of the spring is the indicated value of the target compensation differential pressure. By providing the operating means for adjusting the preset force of the spring, the target compensation differential pressure can be easily adjusted.

The target compensation differential pressure indicating means may be constituted by a hydraulic pressure generating means. In this case, the generated hydraulic pressure becomes an instruction value of the target compensation differential pressure. Also, by introducing a variable pressure into the pressure receiving chamber, the target compensation differential pressure can be similarly adjusted. Further, in this case, since the pressure can be easily adjusted by using an electromagnetic proportional pressure reducing valve or the like, the flow rate passing through the main variable throttle can be more finely controlled by controlling the differential pressure across the main variable throttle.

Further, the above-described configuration is adopted for each of the plurality of directional control valve devices, and the pilot control means is provided with a target value of the differential pressure across the main variable throttle as a target pressure difference between the pump port pressure and the maximum load pressure. By providing target compensation differential pressure indicating means for indicating a corresponding value, control is performed such that the differential pressure across the main variable throttle matches the same target value corresponding to the differential pressure in all of the plurality of direction switching valve devices. . For this reason, if the supply flow rate of the hydraulic power source connected to the pump port is insufficient during the combined operation of simultaneously driving a plurality of actuators, the differential pressure between the supply pressure and the maximum load pressure decreases to reduce the target compensation differential pressure. Are also reduced in common.
A function similar to that of the hydraulic drive device described in Japanese Patent Application Laid-Open No. 706 is obtained, and an appropriate combined operation can be performed.

[0030]

Embodiments of the present invention will be described below with reference to the drawings. First, a first embodiment of the present invention will be described with reference to FIGS. In FIGS. 1 and 2, the hydraulic control valve device of the present embodiment is indicated by reference numeral 100 as a whole. The hydraulic control valve device 100 according to the present embodiment has a housing 1 and a fixed block 2 integrally attached to the housing 1, and a directional switching valve 200 and a seat valve 300 are incorporated in the housing 1. A pilot control valve 400 is incorporated in the fixed block 2. The seat valve 300 performs an auxiliary flow control function as a pressure compensating valve and a load check function in cooperation with the pilot control valve 400.

The directional control valve 200, the seat valve 300, and the pilot control valve 400 are configured as follows. A bore 3 is formed through the housing 1, and the spool valve element 4 of the directional control valve 200 is slidably inserted into the bore 3. In the housing 1, a pump passage 5b communicating with a pump port 5a (see FIG. 2) connected to a hydraulic pressure source (not shown) and a load passage 6 having load ports 6a and 6b connected to an actuator (not shown) are provided.
A, 6B and feeder passages 7A, 7B located between the pump passage 5 and the load passages 6A, 6B are formed.

The bore 3 has annular feeder passages 8A and 8B communicating with the feeder passages 7A and 7B, and load passages 6A and 6B.
B, the annular load passages 9A and 9B communicating with the tank port 85 (see FIG. 2).
0B are formed, and land portions 11A and 12A are formed between the feeder passage 8A and the load passage 9A and between the load passage 9A and the discharge passage 10A, respectively.
B and the load passage 9B and between the load passage 9B and the discharge passage 1
Lands 11B and 12B are respectively formed between the lands 11B and 0B. A load detection passage 12 for detecting a load pressure is formed near the center of the bore 3.
Is formed with a load detection port 13 for taking out the load pressure detected in the load detection passage 12 to the outside.

Notches 14A and 14B are provided on the spool valve body 4.
And notches 15A and 15B. Notch 1
4A forms a meter-in main variable throttle 16A in cooperation with the land 11A. The variable throttle 16A is located between the feeder passage 8A and the load passage 9A, and moves the spool valve body 4 rightward in the figure. The opening area is changed from the fully closed position to a predetermined maximum opening according to the amount. The notch 14B cooperates with the land 11B to form a meter-in main variable aperture 16B.
The variable throttle 16B is located between the feeder passage 8B and the load passage 9B, and has an opening area from the fully closed position to a predetermined maximum opening according to the amount of movement of the spool valve body 4 to the left in the drawing. Change. The notch 15B forms a meter-out main variable throttle 17B in cooperation with the land portion 12B, and the variable throttle 17B is connected to the load passage 9B and the discharge passage 1B.
0B, the opening area is changed from the fully closed position to a predetermined maximum opening according to the amount of movement of the spool valve element 4 to the left in the drawing. The notch 15A cooperates with the land portion 12A to form a meter-out main variable throttle 17A. The variable throttle 17A is located between the load passage 9A and the discharge passage 10A. The opening area is changed from the fully closed position to a predetermined maximum opening degree in accordance with the amount of movement.

The pump passage 5b and the feeder passage 7
A valve body (hereinafter, referred to as a seat valve body) 20 of the seat valve 300 is disposed between the housing A and the base 7B, and the seat valve body 20 slides in a bore 21 orthogonal to the bore 3 formed in the housing 1. It is stored freely. The bore 21 opens into the pump passage 5b as shown in an enlarged view in FIG.
Portion 21a forming an inlet passage 22 of the housing 1, a bore portion 21b having a diameter larger than the bore portion 21a opening on the outer wall surface of the housing 1, and a bore portion having a diameter larger than the bore portion 21a and located adjacent to the bore portion 21b. An annular outlet passage 23 is formed between the bore portions 21a and 21c and communicates with the feeder passages 7A and 7B. Further, the opening end of the bore portion 21b is closed by the fixed block 2, and the hydraulic chamber 24 is formed in the bore portion 21b.
Are formed. A spring 25 for urging the seat valve body 20 in the valve closing direction is disposed in the hydraulic chamber 24. This spring 25 is provided for absorbing vibration, and the urging force of the spring 25 on the seat valve element 20 is so small as to be negligible.

The seat valve element 20 has a seat portion 20a located in the bore portion 21a and a sliding portion 20b located in the bore portions 21b and 21c, and has the above-mentioned diameter of the bore portion 21a and the bore portion 21c. Sliding part 20b corresponding to the size relationship
Has a larger diameter than the sheet portion 20a. The seat portion 20a has a cylindrical shape with a recess 26 formed at the center as shown in the figure, and a plurality of semicircular notches 27 are formed through the cylindrical side wall. A variable aperture 28 is formed in cooperation with the section. The variable throttle 28 is located between the inlet passage 22 and the outlet passage 23, and changes the opening area from the fully closed position to a predetermined maximum opening according to the amount of movement of the seat valve element 20.

On the outer peripheral surface of the sliding portion 20b of the seat valve body 20, a pilot flow groove 31 communicated with the inlet passage 22 via passages 29 and 30 formed inside the seat valve body 20.
Are formed. The pilot flow groove 31 forms a controllable throttle 33 in cooperation with a land 32 formed by a step between the bore 21c and the bore 21b. The controllable throttle 33 is located between the inlet passage 22 and the hydraulic chamber 24, and changes the opening area from the fully closed position to a predetermined maximum opening according to the amount of movement of the seat valve element 20.

Returning to FIG. 1, the fixed block 2 has a hydraulic chamber 2
4 and a passage 36 communicating with the outlet passage 23 via a passage 37 formed in the housing 1, and a pilot control valve 400 is disposed between the passage 35 and the passage 36. . The passages 35 to 37 form a pilot passage connecting the pump passage 5b to the feeder passages 7A and 7B via the inlet passage 22, the passages 29 and 30, and the pilot flow groove 31 (controllable throttle 33).
The control variable throttle 33 changes the pressure in the hydraulic chamber 24 according to the pilot flow rate passing through the pilot passage, and determines the moving amount of the seat valve element 20, that is, the stroke position.

Further, a bottom portion 4 is provided at one end in the fixed block 2.
0a (see FIG. 4), and a bore 40 having the other end opened to the outer surface of the fixed block is formed. The spool valve element of the pilot control valve 400 is slidably formed in the bore 40 (hereinafter, referred to as a pilot valve element). 41 are arranged. As shown, the bore 40 is formed in parallel with the bore 3 of the directional control valve 200, and the pilot valve body 41 is also arranged in parallel with the spool valve body 4.

As shown in the enlarged view of FIG. 4, an annular inlet passage 42 in which the passage 35 opens and an annular outlet passage 43 in which the passage 36 opens are formed near the center of the bore 40. And an outlet land 43 to provide an annular land portion 44. Inlet passage 42 and outlet passage 4
3 also constitutes a part of the pilot passage. The pilot valve element 41 includes a spool portion 41a located on the bore bottom portion 40a side, a spool portion 41b located on the opening end side of the bore 40, and a small diameter portion 41c located near the land portion 44.
And an inclined portion 41d connecting the small diameter portion 41c and the spool portion 41a. The inclined portion 41d forms a pilot variable throttle 45 in cooperation with the land portion 44. The variable throttle 45 is located between the inlet passage 42 and the outlet passage 43 and changes the opening area from a predetermined minimum opening to a predetermined maximum opening according to the amount of movement of the pilot valve element 41.

The open end of the bore 40 is
, And between the screw 46 and the pilot valve element 41, both ends thereof are the pilot valve element 41 and the screw 4.
A spring 47 is provided which abuts the piston 6 and biases the pilot valve body 41 in the valve closing direction. The screw 46 is attached to a screw hole 48 formed at the open end of the bore 40, and a preset force is applied to the spring 47 by the screw 46. This preset force of the spring 47 is applied to the meter-in main variable throttle 16A,
The target value of the differential pressure of 16B, that is, the target compensation differential pressure is indicated, and the spring 47 constitutes a target compensation differential pressure indicating means.

The bottom portion 40a of the bore 40 and the spool portion 41
A pressure receiving chamber 50 is formed between the screw 46 in which the spring 47 is disposed and the spool portion 41b.
Between them, a pressure receiving chamber 51 is formed. A passage 5 that connects the outlet passage 43 to the pressure receiving chamber 50 is provided in the pilot valve body 41.
2, 53 are formed. The pressure in the feeder passages 7A and 7B is introduced into the pressure receiving chamber 50 through the outlet passage 23, the passages 36 and 37 and the outlet passage 43, and the passages 52 and 53, and the pressure is applied to the pilot valve body 41 in the valve closing direction. Is applied to In the fixed block 2, a passage 54 opening to the pressure receiving chamber 51 and passages 55 and 5 formed in the housing 1 are provided.
6, a passage 57 communicating with the load ports 6A, 6B via
A shuttle valve 59 is provided between the passage 54 and the passages 57 and 58 to extract the pressure on the high pressure side of the passages 57 and 58 to the passage 54. The pressure on the high pressure side of the load passages 6A, 6B is introduced into the pressure receiving chamber 51 through the passages 55, 56, the passages 57, 58, the shuttle valve 59, and the passage 54, and the pressure is changed in the valve opening direction of the pilot valve body 41. Is applied to With the configuration of the pressure receiving chambers 50 and 51, the pilot control valve 400 controls the pilot flow rate flowing through the pilot passage according to the differential pressure across the main variable throttles 16A and 16B.

Returning to FIG. 1, when the spool valve element 4 moves to the right in the drawing, the load passage 9A
Axial passage 70A for guiding the load pressure of
And an axial passage 70B and a radial passage 7 that guide the load pressure of the load passage 9B to the load detection passage 12 when the spool valve body 4 moves leftward in the figure.
1B and 72B are formed. Radial passage 71A, 7
1B is formed at a position such that it communicates with the discharge passages 10A and 10B when the spool valve element 4 is at the illustrated neutral position, and the bore 3 is formed when the radial passages 72A and 72B are at the illustrated neutral position. It is formed at a position where it can be closed by the wall surface.

Both ends of the spool valve element 4 project from the end face of the housing 1. The left end of the spool valve element 4 is connected to an operation lever (not shown) via a plug 75, and the right end of the spool valve element 4 is connected to a centering spring mechanism 77 via a plug 76. The centering spring mechanism 77 is a known mechanism that holds the spool valve body 4 at the neutral position when the operation lever is not operated, and one spring 78
And two seats 79 and 80.
The centering spring mechanism 77 is covered by a cover 81 attached to the housing 1.

In the hydraulic control valve device 100 of the present embodiment configured as described above, the seat valve 300 basically operates in the same manner as the seat valve described in JP-A-58-501781. That is, the opening area of the pilot flow groove 31 formed in the seat valve body 20 with respect to the land portion 32 (the opening area of the controllable throttle 33) changes according to the amount of movement (stroke position) of the seat valve body 20, and The amount of movement of the body 20 is determined according to the pilot flow rate passing through the pilot flow groove 31 (the controllable throttle 33). The pilot flow is controlled by the variable throttle 45 of the pilot control valve 400.
Is determined by the opening area. Further, the main flow rate flowing from the inlet chamber 22 to the outlet passage 23 through the variable throttle 28 of the seat valve body 20 is proportional to the pilot flow rate.
Is determined by the opening area. Hereinafter, this will be described in detail.

In FIG. 3, the inlet chamber 2 of the seat valve body 20
2, the effective pressure receiving area of the seat portion 21a located at the position No. 2 is Ap, and the effective pressure receiving area of the annular portion 21c located at the outlet passage 23 is A
z, the effective pressure receiving area of the sliding portion 20b located in the hydraulic chamber 24 is defined as Ac, and the inlet pressure in the inlet chamber 22 (the pump passage 5b
Pp, and the outlet pressure in the outlet passage 23 is P
z, when the control pressure in the hydraulic chamber 24 is Pc, the balance of the pressure receiving areas Ap, Az, and Ac of the seat valve body 20 derives from the following equation: Ac = Az + Ap (1) , Ap · Pp + Az · Pz = Ac · Pc (2) In the formula (1), if Ap / Ac = K, Az / Ac = 1−K is obtained. From the formula (2), Pc = K · Pp + (1−K) · Pz (3) is obtained. Can be Here, assuming that the width of the pilot flow groove 31 is constant at w, the opening area of the control variable throttle 33 at the movement amount x of the seat valve element 20 is wx. Assuming that the pilot flow rate at this time is qs, qs = C1 · wx · (Pp−Pc) ^1/2 (4) where C1: the flow coefficient of the controllable restrictor 33. Is substituted, qs = C1 · w
x {(1-K) (Pp-Pz)} ^1/2 . Therefore, the moving amount x is given by: x = (qs / C1 · w) / {(1−K) (Pp−Pz)} ^1/2 (5) From the expression (5), the difference between the inlet pressure Pp and the outlet pressure Pz is obtained. It can be seen that if the pressure is constant, the movement amount x is determined by qs.

Further, assuming that the opening area of the variable throttle 45 of the pilot control valve 400 is a, the pilot flow rate qs passes through the opening area a, so that qs = C2 · a · (Pc−Pz) ^1/2 . (6) Here, C2: the flow coefficient of the variable throttle 45 By modifying the expression (6), qs = C2 · a · {K · Pp + (1−K) Pz−Pz} ^1/2 = C2 · a · K ^1/2 · (Pp−Pz｝ ^1/2 ) (7) By substituting equation (7) into equation (5), x = (C2 · a / C1 · w) {K / (1−K)} ^1/2 = (C2 / C1 · w) ｛K / (1−K)｝ ^1/2 · a (8) Therefore, as shown in the equation (8), the displacement x of the seat valve element 20 is equal to the pilot passage. Is controlled by the opening area “a” of the variable throttle 45 of the pilot control valve 400 provided in the above.

On the other hand, assuming that the main flow rate flowing from the inlet chamber 22 to the outlet passage 23 through the variable throttle 28 of the seat portion 21A of the seat valve 300 is Qs, and L is the outer diameter of the seat portion 21a, Since the opening area of the variable stop 28 is the product of the outer diameter L and the movement amount x, Qs = C3 · L · x · (Pp−Pz) ^1/2 (9) where C3: the variable stop 28 Flow rate coefficient When equation (5) is substituted into this equation, Qs = {(C3 · L / C1 · w) / (1−K) ^1/2 } · qs (10) where α = (C3 · L / C1 · w) / (1-K) ^1/2
In other words, Qs = α · qs (11) Therefore, it is understood that the main flow rate Qs is proportional to the pilot flow rate qs. Therefore, the total flow rate Qv passing through the seat valve 300 is expressed as follows: Qv = Qs + qs = (1 + α) qs (12)

The seat valve 300 operating as described above further functions as follows in combination with the pilot control valve 400.

A preset force of a spring 47 as target compensation differential pressure indicating means is applied to the pilot valve body 41 of the pilot control valve 400 as an urging force in the valve opening direction, and the pressure in the feeder passage 7A or 7B is adjusted. Pressure receiving chamber 50
Is applied so as to act in the valve closing direction, and the load pressure in the load passage 6A or 6B is applied in the pressure receiving chamber 51 so as to act in the valve opening direction. Therefore, assuming that the load pressure is PL, the pressure conversion value of the preset force of the spring 47 is F, and the pressure receiving areas of the pilot valve element 41 in the pressure receiving chambers 50 and 51 are equal, the feeder passage 7A
Alternatively, since the pressure in 7B is equal to the outlet pressure Pz in the outlet passage 23 of the seat valve 300, the pilot valve body 4
The balance of the force applied to 1 is expressed by PL + F = Pz (13).

By modifying this equation (13), Pz-PL = F (14) If the opening area of the main variable throttle 16A or 16B provided in the spool valve element 4 is A, the seat valve 300 The relationship between the flow rate when the flow rate Qv passing through the main variable throttle 16A or 16B and the differential pressure across the flow path is expressed as follows: Qv = C4 · A · (Pz−PL) ^1/2 (15)

Using equations (12) and (14), (1
By transforming the expression 5), qs = C4 · A / (1 + α) · F ^1/2 (16) is obtained. By transforming equation (15) using equation (14), Qv = C4 · A · F ^1/2 (17) is obtained.

In the above equation (17), the flow rate (the flow rate supplied from the pump passage 5b to the load passage 6A or 6B) Qv passing through the main variable throttle 16A or 16B of the directional control valve 200 is determined by the following equation. This means that it is determined by the preset force F and the opening area A of the main variable throttle 16A or 16B, independently of the supply pressure and the load pressure in the load passage 6A or 6B. Main variable aperture 16 at this time
The target value of the front and rear differential pressure Pz-PL of A or 16B is a value specified by the preset force F from the above equation (14).

Accordingly, the seat valve 300 functions as an auxiliary flow control means for auxiliary controlling the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B. Pz-PL is
The preset force F of the spring 47 is controlled to be equal to the specified target compensation differential pressure regardless of the fluctuation of the load pressure or the supply pressure, and the seat valve 300 performs a pressure compensation function.

When the load increases and the load pressure becomes higher than the supply pressure, and the pressure oil tries to flow backward, the pressure in the hydraulic chamber 24 also increases and the seat valve body 20 is closed. This can be understood from the fact that the pilot flow rate from the inlet chamber 22 to the hydraulic chamber 24 becomes 0, and from the above equation (5), the movement amount of the seat valve body 20 becomes 0. When the movement amount of the seat valve element 20 is 0, the control variable throttle 33 formed by the pilot flow groove 31 is also fully closed. Therefore, backflow of the pressure oil from the load passage 6A or 6B to the pump passage 5b is prevented, and the seat valve 300 performs a load check function.

As described above, the seat valve 300 has an auxiliary flow rate control function and performs a pressure compensation function, and also performs a load check function, whereby the following operation and effect can be obtained.

First, the hydraulic control valve device 1 of the present embodiment
00, the pump passage 5 constituting the main circuit
b. Since only two valves, the seat valve 300 and the directional control valve 200, are arranged in the feeder passages 7A and 7B and the load passages 6A and 6B, the three valves of the directional control valve, the pressure compensating valve and the load check valve are provided. As compared with the conventional valve device arranged in the main circuit, the pressure loss when the pressure oil passes through the main circuit is reduced, and the economical operation of the actuator with small energy loss is enabled.

Second, in a conventional hydraulic control valve device having a pressure compensating valve and a load check valve, it is necessary to form a number of complicated pressure receiving chambers, passages, etc. in a balance piston of the pressure compensating valve. there were. That is, it is necessary to form pressure receiving chambers at both ends of the balance piston independently of the pump port to introduce the inlet pressure and the outlet pressure of the main variable throttle, and to make the target compensation differential pressure of the pressure compensating valve variable. It is necessary to additionally install two pressure receiving chambers. Further, it is necessary to form an inner hole for accommodating the load check valve element of the main circuit inside the balance piston. For this reason, compared with a valve device provided only with a load check valve without a pressure compensation function, the circumference of the balance piston and the balance piston itself become larger, and the valve block becomes longer in the axial direction of the balance piston.
The outer shape of the valve block becomes larger. Also, the production of the valve block becomes complicated.

In this embodiment, a pump passage 5b, an outlet passage 23 and a hydraulic chamber 24 are provided in the housing 1, and the seat valve body 20 is located at a position where a load check valve of a conventional valve device without a pressure compensation function is located. The pilot control valve 400 is arranged by using the fixed block 2 which is arranged and holds the seat valve element 20 separate from the housing 1. For this reason, the axial length L of the seat valve 300 in the portion where the seat valve 300 is located in the housing 1 does not become long unlike a conventional valve device having a pressure compensation function, and the conventional valve device without a pressure compensation function has no pressure compensation function. The dimensions are comparable to those of the valve device, and the size of the housing 1 is reduced. The fixed block 2 is also a pilot valve body 4.
1 is arranged in parallel with the spool valve body 4 to reduce the size. Therefore, the entire valve device becomes compact, which is advantageous in terms of cost, and also eliminates a problem that it is difficult to mount the valve device on a construction machine to be used.

Third, generally, the housing of such a valve device is usually made of a casting, but in the valve device of this embodiment, around the bore 21 where the seat valve element 20 is slidably located. Is simplified, a complicated core configuration can be simplified, and in this aspect also, the configuration can be made advantageous in terms of cost.

A pilot flow groove 31 is formed on the outer peripheral surface of the seat valve body 20 slidable on the housing 1.
A control variable throttle 33 for changing the opening area in accordance with the amount of movement of the seat valve element 20 is provided, and the position of the land 32 of the housing 1 for determining the flow control characteristic is also located on the hydraulic chamber 24 as shown in FIG. Since it is provided by a stepped portion, its processing is also easy.

As described above, according to the hydraulic control valve device 100 of the present embodiment, the pressure loss is reduced while having the pressure compensation function and the load check function, and it is possible to drive the actuator with a small energy loss. Further, the outer shape of the housing is made compact, mounting on a construction machine becomes easy, and further, the manufacturing becomes easier, and the manufacturing cost of the valve device can be reduced.

A second embodiment of the present invention will be described with reference to FIGS. In the drawings, the same members as those shown in FIGS. 1 to 4 are denoted by the same reference numerals. In this embodiment, a load check function with higher liquid tightness is obtained.

Referring to FIGS. 5 and 6, the hydraulic control valve device 101 of the present embodiment has a seat valve 301 and a seat valve 30.
A passage 121 is formed in the first seat valve body 20 instead of the passage 29 shown in FIG. 3, and the passage 121 allows the flow of the pressurized oil from the inlet chamber 22 to the hydraulic chamber 24 and prevents the flow in the opposite direction. A check valve 122 is arranged.

In the first embodiment shown in FIGS. 1 to 4, when the load pressure becomes higher than the supply pressure and the pressure oil tries to flow backward as described above, the seat valve body 20 is moved to the fully closed position. The pilot flow groove 31 is fully closed, and the seat valve 3
00 performs a load check function. However, the sliding portion 20 of the seat valve body 20 in which the pilot flow groove 31 is formed
There is a small gap between the outer peripheral surface of b and the bore 21 to enable sliding, and when the load pressure is extremely high, a small amount of pressure oil flows from the gap through the pilot flow groove 31 into the pump passage 5b. May leak out.

In this embodiment, since the check valve 122 is disposed in the passage 121 in the seat valve body 20 which forms a part of the pilot passage, even a slight leak of pressure oil from such a pilot passage is completely prevented. In addition, an extremely liquid-tight load check function can be obtained.

In this embodiment, the check valve 122 is provided in the seat valve element 20. However, the check valve may be installed anywhere on the pilot passage, for example, by connecting the passage 36 and the passage 37. A check valve may be arranged between the fixed block 2 and the housing 1.

A third embodiment of the present invention will be described with reference to FIGS. In the drawing, the same members as those shown in FIGS. 1 to 4 and FIGS. 5 and 6 are denoted by the same reference numerals. In this embodiment, the preset force of the spring in the pilot control valve can be adjusted from outside.

7 and 8, the hydraulic control valve device 102 of this embodiment has a pilot control valve 401, and the open end of the bore 40 of the pilot control valve 401 is closed by an adjuster screw 130.
0 is attached to a screw hole 48 formed at the opening end of the bore 40. Also, the adjuster screw 130
An operation unit 131 for inserting a hexagon wrench and rotating the hex wrench is integrally provided on the head of the camera. The first screw is provided between the adjuster screw 130 and the pilot valve element 41.
In the same manner as in the embodiment, a spring 47 whose both ends are in contact with the pilot valve body 41 and the screw 130 is arranged, and the preset force of this spring is applied as a biasing force in the valve closing direction of the pilot valve body 41.

In this embodiment, when the operating portion 131 is rotated, the insertion depth of the adjuster screw 130 changes, and the preset force of the spring 47 changes accordingly. As described above, the preset force of the spring 47 indicates the target value (target compensation differential pressure) of the differential pressure across the main variable throttles 16A and 16B of the directional control valve 200, and the main variable throttle 16A,
The pressure compensating characteristic of the seat valve 301 for controlling the flow rate of 16B is set. For this reason, the adjuster screw 13
0, the target compensation differential pressure is adjusted, the pressure compensation characteristic of the seat valve 301 is adjusted, and the hydraulic control valve device 1 is adjusted.
02 can be adjusted.

Therefore, according to the present embodiment, optimal pressure compensation characteristics and flow characteristics are set in accordance with the type of actuator driven by the hydraulic control valve device 102 and the type of load thereof, and the operability is improved. can do.

A fourth embodiment of the present invention will be described with reference to FIGS. In the drawing, the same members as those shown in FIGS. 1 to 4 and FIGS. 5 and 6 are denoted by the same reference numerals. In the present embodiment, a hydraulic pressure generating means is provided instead of a spring as a target compensation differential pressure indicating means of the pilot control valve, and the target compensation differential pressure can be adjusted by making the pressure introduced thereto variable. .

9 and 10, the hydraulic control valve device 103 of this embodiment has a pilot control valve 403, and the pilot control valve 403 is configured as follows. The fixed block 2 has a bottom 140a at one end,
A bore 140 having the other end opened to the outer surface of the fixed block is formed. A spool valve element (hereinafter, referred to as a pilot valve element) 1 of the pilot control valve 403 is slidably formed in the bore 140.
41 are arranged. The bore 140 is also formed in parallel with the bore 3 of the directional control valve 200 (see FIG. 1) as in the previous embodiment, and the pilot valve element 141 is correspondingly formed in parallel with the spool valve element 4 (see FIG. 1). Are located.

An annular pressure receiving chamber 150 is formed adjacent to the bottom 140 a of the bore 140. Near the center of the bore 140, an annular inlet passage 142 opening the passage 35 and an annular outlet passage 143 opening the passage 36. And an annular passage 151 in which the passage 54 is opened.
Between the outlet passage 143 and the outlet passage 143 and the passage 15
1 are provided with annular land portions 144 and 152, respectively. An annular passage 153 is formed at the opening end of the bore 140, and a screw hole 148 is formed at the opening end of the bore 140. A screw 146 is attached to the screw hole 148 to close the open end of the bore 140. Screw 146 and pilot valve element 14
A pressure receiving chamber 154 communicating with the passage 153 is formed between the pressure receiving chamber 154 and the pressure receiving chamber 154.

The pilot valve element 141 has a bore bottom 140
The spool portion 141a located on the side a, the spool portion 141b located on the opening end side of the bore 140, the small diameter portion 141c located near the land 144, and the small diameter portion 1
Inclined portion 14 connecting 41c and spool portion 141a
1d. The inclined portion 141d is the land portion 14.
4 together with a pilot variable throttle 145. The variable throttle 145 has an inlet passage 142 and an outlet passage 1
43, the opening area is changed from a predetermined minimum opening to a predetermined maximum opening according to the amount of movement of the pilot valve element 141.

The pressure receiving chamber 15 extending in the axial direction and opening at the end on the side of the bore bottom 140a is provided inside the pilot valve element 141.
5 and a pressure receiving chamber 156 that extends in the axial direction and opens at the end on the pressure receiving chamber 154 side, and a slidable piston 157 whose one end contacts the bore bottom 140a is inserted into the open end side of the pressure receiving chamber 155. A slidable piston 158 having one end abutting on the screw 148 is provided at the open end of the pressure receiving chamber 156.
Is inserted. In addition, the pilot valve element 141 includes:
A radial passage 159 connecting the pressure receiving chamber 155 to the outlet passage 143 and a radial passage 160 connecting the pressure receiving chamber 156 to the passage 151 are formed. In the pressure receiving chamber 155, the outlet passage 23 and the passages 36 and 37 of the seat valve and the outlet passage 1 are provided.
43 and feeder passage 7A or 7B via passage 159
Is introduced, and the pressure is applied in the valve closing direction of the pilot valve element 141. The passages 55 and 5 are provided in the pressure receiving chamber 156.
6, the pressure on the high pressure side of the load passages 6A, 6B is introduced through the passages 57, 58, the shuttle valve 59, and the passage 54, and the pressure is applied in the valve opening direction of the pilot valve body 141. The pressure receiving chambers 155 and 156 have the same inner diameter, and the piston 1
57, 158 also have the same outer diameter, and the pressure receiving chambers 155, 15
6 and the pressure receiving areas of the pistons 157 and 158 are equalized.

The fixed block 2 has a passage 161 for introducing pressure oil at a constant pressure into the pressure receiving chamber 154, and a passage 1.
A passage 162 for introducing variable-pressure hydraulic oil is formed in 50. The constant pressure introduced into the pressure receiving chamber 154 is applied in the valve opening direction of the pilot valve body 141, and
The pressure introduced into 50 is applied in the valve closing direction of pilot valve element 141.

In the pressure receiving chamber 154, a spring 163 having one end abutting on the pilot valve body 141 and the other end abutting on the screw 146 is arranged. The spring 163 is provided for absorbing vibration. In addition, the urging force of the spring 163 on the pilot valve element 141 is negligibly small.

Therefore, the difference between the oil pressure in the valve opening direction due to the constant pressure introduced into the pressure receiving chamber 154 and the oil pressure due to the variable pressure introduced into the pressure receiving chamber 150 is the target compensation differential pressure in the embodiment shown in FIG. The biasing force acts as a biasing force instead of the preset force of the spring 47 as the indicating means, and the biasing force can be adjusted by controlling the pressure introduced into the pressure receiving chamber 150.

An example of a configuration for generating a constant pressure introduced into the pressure receiving chamber 154 and a variable pressure introduced into the pressure receiving chamber 150 is shown in FIG. In FIG. 9, 500
Is a pilot pump, and a relief valve 501 is connected to a discharge line 507 of the pilot pump to maintain the pressure of the pilot line 502 at a constant pressure Pi. The pilot line 502 is connected to the passage 161 via the pilot line 503, and a constant pressure Pi is introduced into the pressure receiving chamber 154. In addition, the pilot line 50
2 is connected to the primary side of the electromagnetic proportional pressure reducing valve 504, and the secondary side of the electromagnetic proportional pressure reducing valve 504 is connected to the above-mentioned passage 162 via the pilot line 505. The electromagnetic proportional pressure reducing valve 504 is controlled by a control signal from the control device 506, generates a variable pressure Pc according to the control signal, and the variable pressure Pc is introduced into the pressure receiving chamber 150.

In the present embodiment configured as described above, the seat valve 301 functions as follows in combination with the pilot control valve 403.

The pressures and load pressures in the feeder passages 7A and 7B are set to Pz and PL, respectively, as in the first embodiment.
The constant pressure Pi introduced into the pressure receiving chamber 154 and the pressure receiving chamber 150
Assuming that the biasing force due to the difference with the variable pressure Pc introduced into the pilot valve is Fh, the balance of the force applied to the pilot valve element 141 is PL + Fh = Pz, as in the above-described equation (13) according to the first embodiment. (18) is represented by

By transforming this equation (18), Pz-PL = Fh (19) Using the above equation (12) and equation (19), the flow rate when passing through the main variable throttle 16A or 16B (15) expressing the relationship between the pressure and the front-rear differential pressure, qs = C4 · A / (1 + α) · Fh ^1/2 (20) is obtained. By transforming the expression (15), Qv = C4 · A · Fh ^1/2 (21) is obtained. That is, similarly to the first embodiment, the flow rate Qv passing through the main variable throttle 16A or 16B of the directional control valve 200 depends on the urging force Fh and the opening area of the main variable throttle 16A or 16B regardless of the supply pressure and the load pressure. A, and the differential pressure Pz-PL across the main variable throttle at this time is a value indicated by the urging force Fh from the above equation (19).

Therefore, also in this embodiment, the seat valve 301 functions as auxiliary flow rate control means for auxiliary controlling the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B. The 16B front-back differential pressure Pz-PL is controlled to match the target compensation differential pressure indicated by the urging force Fh irrespective of fluctuations in the load pressure or the supply pressure, and the seat valve 301 performs a pressure compensation function.
That is, the seat valve 301 can have a pressure compensation function and a load check function.

In the present embodiment, the urging force Fh can be adjusted by adjusting the pressure Pc, and the pressure Pc can be easily adjusted by using the control device 506, the electromagnetic proportional pressure reducing valve 504, and the like. It can be performed with good controllability. Therefore, more precise adjustment of the target compensation differential pressure is possible, and thereby the directional control valve 20 can be more appropriately adjusted.
Flow rate Qv passing through the main variable throttle 16A or 16B of 0
And the operability of the actuator can be further improved.

A fifth embodiment of the present invention will be described with reference to FIGS. In the drawing, the same members as those shown in FIGS. 1 to 4, 5, 6, and 9 are denoted by the same reference numerals. In this embodiment, a means for applying an urging force based on the pressure difference between the pump discharge pressure and the maximum load pressure is provided as target compensation differential pressure indicating means of the pilot control valve.

In FIGS. 11 and 12, the hydraulic control valve device 104 of this embodiment has a first directional switching valve device 104A and a second directional switching valve device 104B, and the first directional switching valve device 104A is Direction switching valve 204, seat valve 301
And a pilot control valve 404.

That is, in FIG. 11, a bore 220 is formed through the housing 1, and the spool valve element 221 of the directional control valve 204 is slidably inserted into the bore 220. In the housing 1, load passages 6A and 6B having load ports 6a and 6b connected to an actuator (not shown), a pump passage 222 branched from the pump passage 222b, the pump passage 222 and the load passages 6A and 6 are provided.
B and feeder passages 7A and 7B located between them.

In the bore 220, as in the embodiment shown in FIG. 1, annular feeder passages 8A and 8B, annular load passages 9A and 9B, and annular discharge passages 10A and 10B are formed. Lands 11A, 11B and 12 between
A and 12B are formed respectively. Also, bore 22
0, the pump passage 222b is formed as an annular passage in the center of the pump port 222b.
2a is connected to the hydraulic pump 600 (FIG. 1).
2).

The notch 224A,
224B and notches 225A, 225B are formed. The notch 224A cooperates with the land 11A to form a meter-in main variable throttle 226A. The variable throttle 226A is located between the feeder passage 8A and the load passage 9A. The opening area is changed from the fully closed position to a predetermined maximum opening according to the amount of movement. The notch 224B cooperates with the land portion 11B to form a meter-in main variable throttle 226B. The variable throttle 226B is located between the feeder passage 8B and the load passage 9B. The opening area is changed from the fully closed position to a predetermined maximum opening according to the amount of movement. The notch 225B cooperates with the land portion 12B to form a meter-out main variable throttle 227B. The variable throttle 227B is located between the load passage 9B and the discharge passage 10B on the right side in the drawing of the spool valve element 221. The opening area is changed from the fully closed position to a predetermined maximum opening according to the amount of movement.
The notch 225A forms a meter-out main variable aperture 227A in cooperation with the land 12A.
27A is located between the load passage 9A and the discharge passage 10A, and changes the opening area from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the spool valve element 221 to the left in the drawing.

The pump passage 222 and the feeder passage 7
Between A and 7B, a valve body (hereinafter, appropriately referred to as a seat valve body) 20 of the seat valve 301 is disposed. The configuration of the seat valve 301 is the same as that of the second embodiment shown in FIG. 5, and a description thereof will be omitted.

The land portions 11A and 11B have annular load detection chambers 230A and 230A for detecting the load pressure.
B is formed, and the load detection chamber 230A,
Load detection passages 231A and 231B connected to 230B are formed. The load detection chamber 230A is provided with the spool valve 2
The load detection chamber 230B is provided at a position where the load pressure of the load passage 9A is taken out when 21 moves rightward in the figure.
It is provided at a position where the load pressure of the load passage 9B is taken out when the spool valve element 221 moves to the left in the figure. Further, passages 232A, 233 are provided in the spool valve body 221.
A, 234A are formed, and the load detection chamber 230A and the load detection passage 231A are connected to these passages 232A, 233A, 234A when the spool valve element 221 returns to the neutral position.
To reduce the detected load pressure to the tank pressure. Similar passages are provided in the spool valve body 221 for the load detection chamber 230B and the load detection passage 231B. As described above, by reducing the load pressure detected when the spool valve element 221 is in the neutral state, it is possible to prevent the discharge pressure of the hydraulic pump during the neutral state from increasing needlessly in the load sensing type hydraulic drive device.

On the other hand, a pilot control valve 404 is incorporated in the fixed block 2. This pilot control valve 4
The configuration of 04 is similar to that of the embodiment shown in FIG. 9, and the configuration is enlarged and shown in FIG. In the figure, the same members as those shown in FIG. 9 are denoted by the same reference numerals.

In FIG. 13, a bore 140 is formed in the fixed block 2, and a spool valve element (hereinafter, referred to as a pilot valve element) 141 of a pilot control valve 404 is slidably disposed in the bore 140.

As in the embodiment shown in FIG. 9, an annular pressure receiving chamber 150, an annular inlet passage 142, an annular outlet passage 143, an annular passage 151, and an annular passage 1 are provided in the bore 140.
53 and a screw hole 148 are formed.
A screw 146 is attached to 8, and the open end of the bore 140 is closed. The pressure receiving chamber 15 communicating with the passage 153 between the screw 146 and the pilot valve element 141 is also provided.
4 are formed, and a weak spring 163 for preventing vibration is arranged in the pressure receiving chamber 154. Inlet passage 142 and outlet passage 1
A pilot variable throttle 145 is formed between the land portion 144 formed between the pilot valve member 43 and the inclined portion 141d of the pilot valve element 141. Further, another annular passage 239 is formed between the pressure receiving chamber 150 and the inlet passage 142.

Inside the pilot valve element 141, pistons 157, 158 extending in the axial direction and slidable toward the open end.
Are formed, and the pressure receiving chambers 240 and 241 are respectively provided with radial passages 242 and 243.
Through the passages 239, 151.

In the fixed block 2, a passage 251 for communicating the pressure receiving chamber 150 with the feeder passages 7A and 7B via a passage 250 formed in the housing 1, and a passage 252 for communicating the passage 153 with the load detection passages 231A and 231B.
Is formed, and the pressure in the feeder passage 7A or 7B is introduced into the pressure receiving chamber 150 through the passages 250 and 251.
The pressure is applied in the valve closing direction of the pilot valve element 141, and the pressure receiving chamber 154 has load detection chambers 230 A, 230 B,
The pressure in the load passage 6A or 6B is introduced through the passages 231A and 231B, the passage 252, and the passage 153, and the pressure is applied in the valve opening direction of the pilot valve body 141.

The fixed block 2 includes passages 253 and 254 for connecting the passage 151 to the pump passage 222,
Passage 255 communicating with load detection passages 231A and 231B
And passages 256 and 257 communicating with similar load detection passages of a directional control valve (not shown), and passages 258, 259 and 260 communicating with the passage 239 are formed between the passage 260 and the passages 255 and 256. A shuttle valve 261 for extracting the pressure on the high pressure side of the passages 255 and 256 to the passage 260 is provided. These passages 253, 2
The supply pressure of the pump port, that is, the discharge pressure of the hydraulic pump is introduced through the passage 54 and the passages 151 and 243, and the pressure is applied in the valve opening direction of the pilot valve body 141. The maximum load pressure of a plurality of actuators is introduced into the pressure receiving chamber 240 through the passages 255, 256, 257, the shuttle valve 261, the passages 258, 259, 260, and the passages 239, 242. In the valve closing direction.

The fixed block 2 is further provided with a load detection port 262 which communicates with the passage 259 and takes out the maximum load pressure to the outside.

The second directional control valve device 104B is shown in FIG. This second directional control valve device 104
B has substantially the same structure as the first directional control valve device 104A, and the same members as those of the first directional control valve device 104A are denoted by the same reference numerals and description thereof is omitted.

The hydraulic control valve device 1 configured as described above
FIG. 12 shows a circuit configuration of a hydraulic drive device in which No. 04 is used. In FIG. 12, reference numeral 600 denotes a variable displacement hydraulic pump whose displacement is controlled by a load sensing type regulator 601. The discharge pipeline 602 of the hydraulic pump 600 is connected to the pump port 222a of the hydraulic control valve device 104. 603,6
Reference numeral 04 denotes a hydraulic actuator, and the load ports 6a and 6b of the first direction switching valve device 104A are connected to the first actuator 603 through actuator lines 605A and 605B.
And the second actuator 604 is connected to the second
Are connected to the load ports 6a, 6b of the directional switching valve device 104B via actuator lines 606A, 606B. Further, the first and second directional switching valve devices 104
The tank ports 85 of A and 104B are connected to the tank 607. With this configuration, the discharge pressure of the hydraulic pump 600 is introduced into the passage 254, the load pressure on the high pressure side of the hydraulic actuators 603, 604 is introduced as the maximum load pressure into the passage 260, and the pressure receiving chamber 2
41 and 240.

The discharge pressure of the hydraulic pump 600 is introduced into the regulator 601 through the pilot line 608, and the maximum load pressure is guided through the pilot line 609 connected to the load detection port 262. As is well known, the regulator 601 controls the displacement of the hydraulic pump 600 based on the pump discharge pressure and the maximum load pressure such that their differential pressures maintain a predetermined value.

Therefore, in the pilot control valve 404, the biasing force due to the pressure difference between the pump discharge pressure introduced into the pressure receiving chamber 241 and the maximum load pressure introduced into the pressure receiving chamber 240 becomes the target compensation in the embodiment shown in FIG. Acting instead of the preset force of the spring 47 as the differential pressure indicating means, the seat valve 301 can be provided with a pressure compensation function and a load check function as in the first embodiment.

That is, the pressure receiving chambers 150 and 154 have the same pressure receiving area, the pressure receiving chambers 240 and 241 have the same pressure receiving area, and the pressure in the feeder passages 7A and 7B and the load pressure in the feeder passages 7A and 7B are respectively similar to the first embodiment. Pz, PL, the discharge pressure of the hydraulic pump 600 is Pp, and the maximum load pressure is PLSma.
Assuming that x, the balance of the force applied to the pilot valve element 141 is expressed by Pp + PL = Pz + PLSmax (22) as in the above-described equation (13) according to the first embodiment.

By modifying the equation (22), Pz-PL = Pp-PLSmax = Fd (23) Therefore, the differential pressure between the pump discharge pressure and the maximum load pressure becomes the urging force Fd of the flow rate indicating means.

The above equation (12) and this equation (23)
Using the above equation, the above-mentioned (1) representing the relationship between the flow rate when passing through the main variable throttle 226A or 226B and the differential pressure before and after.
By transforming equation 5), the pilot flow rate qs and the urging force Fd
Is expressed as: qs = C4 · A / (1 + α) · Fd ^1/2 (24) By transforming equation (15) using equation (19), the flow rate Q supplied from the pump port to the load port
v is represented by Qv = C4 · A · Fd ^1/2 (24) That is, similarly to the first embodiment, the flow rate Qv passing through the main variable throttle 226A or 226B of the directional control valve 204 depends on the urging force Fd and the opening area of the main variable throttle 226A or 226B regardless of the supply pressure and the load pressure. A
At this time, the differential pressure Pz-P before and after the main variable throttle.
L is a value indicated by the urging force Fd from the above equation (22).

Therefore, also in this embodiment, the seat valve 301 functions as auxiliary flow rate control means for auxiliary controlling the flow rate of the pressure oil supplied to the main variable throttle 226A or 226B. 226
The pressure difference Pz-PL before and after B is controlled to match the target compensation pressure difference indicated by the urging force Fd irrespective of the change in the load pressure or the supply pressure, and the seat valve 301 performs a pressure compensation function. That is, the seat valve 301 can have a pressure compensation function and a load check function.

In this embodiment, the target value (target compensation differential pressure) of the front and rear differential pressures of the main variable throttles of the first and second direction switching valve devices 104A and 104B is controlled by the hydraulic pump 600 under load sensing control. Of the hydraulic pump 60 during the combined driving of the hydraulic actuators 603 and 604.
When the discharge flow rate of 0 is insufficient, the differential pressure between the pump discharge pressure and the maximum load pressure decreases, and the target value of the differential pressure across the main variable throttle also decreases in common for the two directional control valves. Therefore, as in the case of the hydraulic drive system described in Japanese Patent Application Laid-Open No. 60-11706, the problem that a large amount of pressure oil is supplied to the light-load-side actuator and the heavy-load-side actuator is not driven is solved. A composite operation becomes possible.

[0108]

According to the present invention, the seat valve has a pressure compensation function and a load check function, and two valves, a seat valve and a main variable throttle, are provided in the pump port, the feeder passage and the load port constituting the main circuit. Since it is only arranged, the pressure loss when the pressure oil passes through the main circuit is reduced, and the energy loss can be reduced.

Further, since the seat valve is provided with a pressure compensation function and a load check function in combination with the pilot control means, the structure around the seat valve body is simplified, the housing is made compact, and the housing is easily manufactured. become.

Further, since the pilot control means is disposed by using the fixed block for holding the seat valve element, the installation of the pilot control means becomes easy, and the pilot valve element of the pilot control valve is arranged in parallel with the spool valve element. The arrangement also makes the fixed block compact.

Since the check valve is arranged in the pilot passage,
A more liquid-tight load check function can be obtained.

Since the target compensation differential pressure indicating means is constituted by a spring, the structure is simple, and the preset force of the spring can be adjusted, so that the adjustment of the target compensation differential pressure becomes easy.

Since the target compensation differential pressure indicating means is constituted by the hydraulic pressure generating means and a variable pressure is introduced, the adjustment of the target compensation differential pressure is further facilitated, and delicate flow control becomes possible.

Further, as the target compensation differential pressure indicating means of each of the plurality of direction switching valve devices, an urging means for applying an urging force corresponding to a differential pressure between the supply pressure of the pump port and the maximum load pressure to the pilot valve body. , An appropriate composite operation can be performed.

[Brief description of the drawings]

FIG. 1 is a sectional view of a hydraulic control valve device according to a first embodiment of the present invention.

FIG. 2 is a circuit diagram of the hydraulic control valve device shown in FIG.

FIG. 3 is an enlarged view of a seat valve in the hydraulic control valve device shown in FIG.

FIG. 4 is an enlarged view of a pilot control valve in the hydraulic control valve device shown in FIG.

FIG. 5 is a sectional view of a hydraulic control valve device according to a second embodiment of the present invention.

6 is a circuit diagram of the hydraulic control valve device shown in FIG.

FIG. 7 is a sectional view of a hydraulic control valve device according to a third embodiment of the present invention.

8 is a circuit diagram of the hydraulic control valve device shown in FIG.

FIG. 9 is a cross-sectional view of a pilot control valve portion of a hydraulic control valve device according to a fourth embodiment of the present invention and a circuit diagram of a related circuit configuration thereof.

FIG. 10 is a circuit diagram of the hydraulic control valve device shown in FIG.

FIG. 11 is a sectional view of a hydraulic control valve device according to a fifth embodiment of the present invention.

FIG. 12 is a circuit diagram of the hydraulic control valve device shown in FIG.

FIG. 13 is an enlarged view of a pilot control valve in the hydraulic control valve device shown in FIG.

[Explanation of symbols]

REFERENCE SIGNS LIST 100 Hydraulic control valve device 200 Direction switching valve 300 Seat valve 400 Pilot control valve 1 Housing 2 Fixed block 3 Bore 4 Spool valve element 5a Pump port 5 Pump passage 6A, 6B Load passage 6a, 6b Load port 7A, 7B Feeder passage 16A, 16B Main variable throttle 20 Seat valve element 21 Bore 31 Pilot flow groove 33 Control variable throttle 35, 36, 37 Pilot passage 40 Bore 41 Pilot valve element 45 Pilot variable throttle 47 Spring (target compensation differential pressure indicating means) 50, 51 Pressure receiving chamber (Biasing means) 301 Seat valve 122 Check valve 401 Pilot control valve 130 Adjuster screw 131 Operating section 403 Pilot control valve 150,154 Pressure receiving chamber (Target compensation differential pressure indicating means; Oil pressure generating means) 104A, 104B Direction switching valve Dress 204 directional control valve 404 pilot control valve 221 spool valve body 222 pump port 226A, 226B main variable throttle 240 and 241 receiving chamber (target compensation differential pressure command means; biasing means) 261 shuttle valve (maximum load pressure detecting means)

Continuation of front page (72) Inventor Masami Ochiai 650, Kandamachi, Tsuchiura-shi, Ibaraki Pref. Hitachi Construction Machinery Co., Ltd. Tsuchiura Works (56) References JP-A-4-8902 (JP, A) JP-A-63-235706 ( JP, A) International Publication No. 92/1163 (WO, A1) (58) Fields investigated (Int. Cl. ⁷ , DB name) F15B 11/00-11/22

Claims (11)

(57) [Claims] A housing, a pump port formed in the housing and connected to a hydraulic pressure source, at least one spool valve body slidably disposed in the housing, and formed in the housing; A load passage having a feeder passage communicating with a pump port and a load port connected to an actuator; and a load passage located between the feeder passage and the load passage, wherein an opening area is changed according to an amount of movement of the spool valve element. A hydraulic control valve device having a main variable throttle, wherein auxiliary control of a flow rate of pressure oil supplied to the main variable throttle from the pump port via the feeder passage according to a differential pressure across the main variable throttle. Auxiliary flow rate control means, wherein the auxiliary flow rate control means comprises: (a) an inside of the housing between the pump port and the feeder passage; A slidably disposed seat valve element, a control variable throttle formed on the seat valve element, and configured to change an opening area according to an amount of movement of the seat valve element, and the pump port via the control variable throttle. A seat valve having a pilot passage communicating with the feeder passage, and determining a movement amount of the seat valve element according to a pilot flow rate passing through the pilot passage; and (b) being disposed in the pilot passage and being provided with the main variable throttle. And a pilot control means for controlling the pilot flow rate according to the pressure difference between the front and rear. 2. The hydraulic control valve device according to claim 1, further comprising: a fixed block for holding the seat valve body in the housing via a spring, wherein the pilot control means is a pilot built in the fixed block. A hydraulic control valve device, which is a control valve. 3. The hydraulic control valve device according to claim 2, wherein the pilot control valve includes a pilot valve disposed parallel to the spool valve. 4. The hydraulic control valve device according to claim 1, wherein the directional control valve device further includes a check valve disposed in the pilot passage, for preventing a backflow of the pressure oil in the pilot passage. A hydraulic control valve device characterized by the above-mentioned. 5. The hydraulic control valve device according to claim 1, wherein the pilot control means includes a slidable pilot valve element and an urging force corresponding to the front-rear differential pressure of the main variable throttle. Urging means for applying in the valve closing direction, and applying a predetermined urging force to the pilot valve body in the valve opening direction,
And a target compensation differential pressure indicating means for indicating a target value of the differential pressure across the main variable throttle. 6. The hydraulic control valve device according to claim 5, wherein the target compensation differential pressure indicating means includes a spring acting on the pilot valve body, and a preset force of the spring is used as the predetermined urging force in a valve opening direction. A hydraulic control valve device characterized by being applied to a hydraulic control valve device. 7. The hydraulic control valve device according to claim 6, wherein said pilot control means further includes an operation means for adjusting a preset force of said spring. 8. The hydraulic control valve device according to claim 5, wherein said target compensation differential pressure indicating means includes a hydraulic pressure generating means for generating a predetermined hydraulic pressure as said predetermined urging force. Control valve device. 9. The hydraulic control valve device according to claim 8, wherein said hydraulic pressure generating means includes a pressure receiving chamber into which a variable pressure is introduced. 10. A housing, a pump port formed in the housing and connected to a hydraulic pressure source, and a plurality of directional control valves incorporated in the housing and controlling flow rates of pressure oil supplied to a plurality of actuators, respectively. A plurality of directional control valve devices, each of which includes a spool valve body slidably disposed in the housing, a feeder passage formed in the housing and communicating with the pump port, and the actuator. A hydraulic control valve device, comprising: a load passage having a load port to be connected; and a main variable throttle located between the feeder passage and the load passage, the opening being changed in accordance with the amount of movement of the spool valve element. In the plurality of directional control valve devices, each of the plurality of directional control valve devices may be connected to the feeder passage from the pump port in accordance with a pressure difference across the main variable throttle. Auxiliary flow rate control means for auxiliaryly controlling the flow rate of the pressure oil supplied to the main variable throttle through the auxiliary flow rate control means, wherein: (a) the auxiliary flow rate control means is provided between the pump port and the feeder passage. A seat valve body slidably disposed in the housing, a controllable throttle formed on the seat valve body and changing an opening area in accordance with an amount of movement of the seat valve body, and the controllable control of the pump port; A seat valve having a pilot passage communicating with the feeder passage via a throttle, and determining a moving amount of the seat valve element by a pilot flow rate passing through the pilot passage; and (b) being disposed in the pilot passage, Pilot control means for controlling the pilot flow rate in accordance with the pressure difference between the front and rear of the main variable throttle. The pilot control means further includes means for detecting pressure as a maximum load pressure, wherein the pilot control means sets a value corresponding to a differential pressure between the pump port pressure and the maximum load pressure as a target value of the differential pressure across the main variable throttle. A hydraulic control valve device comprising target compensation differential pressure instruction means for instructing. 11. The hydraulic control valve device according to claim 10, wherein the pilot control means of each of the plurality of directional control valve devices includes a slidable pilot valve element and the main variable throttle provided on the pilot valve element. First urging means for applying an urging force in the valve closing direction in accordance with the differential pressure between the front and rear of the pump valve. The target compensation differential pressure indicating means includes a pressure of the pump port and the maximum load applied to the pilot valve element. A hydraulic control valve device, comprising: a second urging unit that applies an urging force in a valve opening direction according to a pressure difference from the pressure.