JP3144915B2 - 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 hydraulic control valve device having a center bypass type directional control valve provided with a pressure compensating function.

[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. There is a hydraulic drive device provided with a center bypass type directional switching valve described in JP-A-5928 and JP-B-62-38496. In this hydraulic drive device, pressure oil discharged from a hydraulic pump is supplied to a plurality of actuators via a hydraulic control valve device including a plurality of center bypass type directional switching valves. Has a pump port connected to the hydraulic pump, a load port connected to the actuator, a tank port connected to the tank, and a center bypass passage connecting the pump port and the tank port. The bypass passages are connected in series in the hydraulic circuit.

[0003] In the hydraulic drive device constructed as described above, the opening area of the bleed-off variable throttle provided in the center bypass passage gradually decreases in accordance with the amount of movement of the spool valve element of the direction switching valve. The opening area of the main variable throttle provided between the pump port and the load port gradually increases. As a result, the flow rate flowing from the hydraulic pump to the tank through the center bypass line decreases, and correspondingly, the pressure oil from the hydraulic pump is supplied to the actuator through the main variable throttle, and the actuator is driven. You.

On the other hand, as another hydraulic drive device for driving a plurality of actuators with pressure oil discharged from one hydraulic pump, for example, a load sensing type hydraulic drive device disclosed in Japanese Patent Application Laid-Open No. Sho 60-11706 is known. In this hydraulic drive device, a closed center type directional control valve, a pressure compensating valve, and a load check valve are used in combination. When a hydraulic drive device is configured by combining the directional control valve, the pressure compensating valve, and the load check valve as described above, these three valves are incorporated in one block for the purpose of reducing the number of pipes and achieving compactness. At 1
It is considered that the valve device is configured as one valve device. As an example, there is a valve device described in JP-A-02-134402.

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, feeder port, and actuator port are formed in the valve block as part of the main circuit that supplies hydraulic oil from the hydraulic pump to the actuator.The spool of the directional control valve connects 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]

Hitherto, as a hydraulic control valve device for construction machinery, a hydraulic control valve device having a spool type directional switching valve using a spool valve element has been used for many years.
The spool type directional control valve has high reliability and is easy to design because of its many years of use. The center bypass type directional switching valve also uses a spool valve body, and is classified as a spool type directional switching valve. However, the control valve device provided with the conventional center bypass type directional control valve has the following problems.

In the above-mentioned conventional hydraulic control valve device having a center bypass type directional control valve, the plurality of directional control valves are connected in parallel. When the operation shifts to the combined operation in which the pressure oils are simultaneously driven, the supply flow rate of the pressure oil changes suddenly, and operability may be impaired. That is, the load pressure of the actuator in the single operation is relatively low, and the operation amount of the direction switching valve (stroke amount of the spool valve element) is relatively small, and the actuator operates at a very low speed.
If the load pressure of the other actuator is higher than the load pressure of the previous actuator when the operation shifts to the combined operation, the actuators with the lower load pressure are connected because the pump passages of the directional control valves of both actuators are connected in parallel. More pressure oil flows in, and an unexpected increase in the speed of the actuator is started by the combined operation. In a hydraulic excavator or a crane, for example, the behavior of a low load side actuator is a swing motor, and the high load side actuator is a boom cylinder. This occurs when the directional control valve is operated to raise the boom in the middle. In this case, the supply flow rate to the swing motor suddenly increases, and the swing speed suddenly changes, resulting in a very dangerous state.

In order to solve such a problem, it is conceivable to arrange a pressure compensating valve in a pump passage like a closed center type directional switching valve even in a center bypass type directional switching valve. However, when the pressure compensating valve is provided in the body of the actual directional control valve, the valve spool of the pressure compensating valve and the bore for slidably housing the valve spool may be provided in the housing of the directional control valve. is necessary. However, the addition of a new configuration to the housing of the directional control valve in the manner described in Japanese Patent Laid-Open No.
As seen from the valve device described in Japanese Patent No. 4402, the structure becomes complicated and the size becomes large, which is extremely disadvantageous in terms of cost and the like.

That is, in the valve device described in Japanese Patent Application Laid-Open No. 02-134402, a pressure receiving chamber is formed independently of a pump port at both ends of a balance piston of a pressure compensating valve, and the inlet pressure and the outlet pressure of the main variable throttle are controlled. In order to make the target compensation differential pressure of the pressure compensating valve variable, it is necessary to additionally provide 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 center bypass type directional switching valve which is a spool type is not considered at all.

A first object of the present invention is to provide a hydraulic control valve device provided with a center bypass type directional switching valve device by adding a pressure compensating function to reduce the supply flow rate when shifting from a single operation to a combined operation. An object of the present invention is to provide a hydraulic control valve device that prevents sudden changes and has excellent operability.

A second object of the present invention is to add a pressure compensating function and a load check function to a center bypass type directional switching valve device without complicating the structure of a housing, and to provide a compact and easy-to-manufacture hydraulic control valve. It is to provide a device.

[0014]

In order to achieve the first and second objects, according to the present invention, a housing is connected to a pump port and a tank formed in the housing and connected to a hydraulic pressure source. A first spool valve body slidably disposed in the housing, a pump passage formed in the housing and communicating with the pump port, a feeder passage communicating with the pump passage, A load passage having a load port connected to a first actuator and a discharge passage communicating with the tank port; a first center bypass passage formed in the housing and communicating the pump port with the tank port; A main spool that is located between the feeder passage and the load passage and that changes an opening area according to an amount of movement of the first spool valve body; A first direction switching valve of a center bypass type having a variable throttle and a main variable throttle for bleed-off which is located on the center bypass passage and changes an opening area according to an amount of movement of the first spool valve element. A second spool valve body slidably disposed in the housing, and a second center bypass passage connected in series with the first center bypass passage;
A center bypass type second control device that controls the flow rate of pressure oil supplied from the pump port to the second actuator.
A hydraulic control valve device comprising:
The first directional control valve device assists in assisting control of a flow rate of pressure oil supplied to the main variable throttle from the pump passage via the feeder passage in accordance with a differential pressure across the main variable throttle. A flow control means, the auxiliary flow control means comprising: (a) a seat valve body slidably disposed in the housing between the pump passage and the feeder passage;
A control variable throttle formed in the seat valve body and changing an opening area in accordance with a movement amount of the seat valve body, and a pilot passage connecting the pump passage to the feeder passage through the control variable throttle. A seat valve for determining the amount of movement of the seat valve body by a pilot flow rate passing through the pilot passage; and (b) being disposed in the pilot passage and controlling the pilot flow rate according to a pressure difference across the main variable throttle. And a pilot control means for controlling the hydraulic pressure.

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 first direction switching 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 front-rear differential pressure of 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.

[0020]

In the hydraulic control valve device of the present invention constructed as described above, when the spool valve element of the first directional control valve device is operated from the neutral position, the main variable throttle opens, and most of the pressure oil in the pump passage is removed. At the same time, the main oil flows through the seat valve and flows out to the feeder passage, and a part of the pressure oil in the pump passage flows as the pilot flow through the pilot passage of the seat valve 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. Therefore, when the pilot flow rate is controlled in accordance with the differential pressure across the main variable throttle, the amount of movement of the seat valve body is adjusted accordingly, 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 auxiliary controlled, and the differential pressure across the main variable throttle is 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 added.

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 also 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. As described above, the first directional control valve device is provided with the pressure compensation function in addition to the load check function, so that the first directional control valve device can be changed from a single operation for driving one of the plurality of actuators to a combined operation for simultaneously driving the other actuators. In the case of the shift, the sudden change of the flow rate of the supply of the pressure oil to the previous actuator is prevented, and the operability is improved.

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 passage, the feeder passage and the load passage constituting the main circuit. The pressure loss is equivalent to that of the hydraulic control valve device including the center bypass type directional control valve without compensation function and the load check valve, and the pressure compensation function is added without increasing the pressure loss.

Also, in the hydraulic control valve device of the present invention, the seat valve is located at the position in the housing where the load check valve of the conventional hydraulic control valve device having the closed center type directional switching valve without pressure compensation function was provided. Arrange the seat valve,
Pilot control means is provided separately from the seat valve, and the pilot flow rate for determining the moving amount of the seat valve body 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.

[0029]

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. 1 and 2, the hydraulic control valve device of the present embodiment is generally indicated by reference numeral 100. The hydraulic control valve device 100 of this embodiment drives a hydraulic actuator 701 as shown in FIG. First directional switching valve device 1
00A, second directional switching valve device 100B for driving hydraulic actuator 702, hydraulic actuator 70
3 has a third direction switching valve device 100C for driving the third direction switching valve device 100C.

Further, the hydraulic control valve device 100 has a housing 1 common to the first to third direction switching valve devices, and a fixed block 2 integrally attached to the housing 1, and has a first direction. The switching valve device 100A is a center bypass type directional switching valve 20 incorporated in the housing 1.
It has a 0A and seat valve 300 and a pilot control valve 400 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 200A, the seat valve 300, and the pilot control valve 400 are each configured as follows. A bore 3 is formed through the housing 1.
The spool valve element 4A of the direction switching valve 200A is slidably inserted into the bore 3. Also, a pump passage 5b having a pump port 5a (see FIG. 2) connected to the hydraulic pump 700 in the housing 1, a pump passage 5 branched from the pump passage 5b, and a hydraulic actuator 701 (see FIG. 2) are connected. Load passages 6A and 6B having load ports 6a and 6b to be connected, a pump passage 5 and load passages 6A and 6B.
B and feeder passages 7A and 7B located between them.

Near the center of the bore 3, there is a pump port 5a.
An annular entry-side center bypass passage 750 communicating with the
Annular outlet center bypass passages 751A and 751B communicating with the outlet side center bypass passage 751 (see FIG. 2).
Are formed, and land portions 752A and 752B are formed between the entrance-side center bypass passage 750 and the exit-side center bypass passages 751A and 751B, respectively.
The bore 3 has annular feeder passages 8A and 8B communicating with the feeder passages 7A and 7B, annular load passages 9A and 9B communicating with the load passages 6A and 6B, and a tank port 85.
Annular discharge passages 10A and 10B communicating with (see FIG. 2)
Are formed between the feeder passage 8A and the load passage 9A and between the load passage 9A and the discharge passage 10A.
1A and 12A are respectively formed between the feeder passage 8B and the load passage 9B, and between the load passage 9B and the discharge passage 10B.
And lands 11B and 12B are formed between them. The tank port 85 is connected to the tank 704.

Notches 753A and 73 are provided on the spool valve body 4A.
53B and a cylindrical portion 755 are formed. Notch 75
3A and the cylindrical portion 755 are the land portions 752A, 752
In cooperation with B, a bleed-off variable throttle 754A is formed, and the variable throttle 754A is provided with an inlet-side center bypass passage 750 and outlet-side center bypass passages 751A, 751B.
The opening area is changed from the fully open position to the fully closed position as shown by PT in FIG. 3 according to the amount of movement (spool stroke) of the spool valve element 4A to the right in the drawing.
The notch 753B and the cylindrical portion 755 are connected to the land portion 752.
B, 752A in cooperation with bleed-off variable aperture 754
B, the variable throttle 754B includes an inlet-side center bypass passage 750 and an outlet-side center bypass passage 751B,
751A, and changes the opening area from the fully open position to the fully closed position as shown by PT in FIG. 3 according to the amount of movement of the spool valve element 4A to the left in the drawing.

A notch 14 is provided on the spool valve element 4A.
A, 14B and notches 15A, 15B are formed. The notch 14A forms a meter-in main variable aperture 16A in cooperation with the land 11A.
A is located between the feeder passage 8A and the load passage 9A,
3 according to the amount of movement of the spool valve element 4A to the right in the drawing.
As shown in -A, the opening area is changed from the fully closed position to a predetermined maximum opening degree. The notch 14B is located at the land 11B.
And a main variable throttle 16B of a meter-in, which is located between the feeder passage 8B and the load passage 9B, as shown in FIG. The opening area is changed from the fully closed position to a predetermined maximum opening according to the leftward movement amount. Notch 15B
Forms a meter-out main variable throttle 17B in cooperation with the land portion 12B, and the variable throttle 17B
B and the discharge passage 10B.
The opening area is changed from the fully closed position to a predetermined maximum opening as shown by BT in FIG. The notch 15A forms a meter-out main variable aperture 17A in cooperation with the land 12A.
7A is located between the load passage 9A and the discharge passage 10A,
3B according to the rightward movement of the spool valve element 4A in the drawing.
The opening area is changed from the fully closed position to a predetermined maximum opening as shown by -T.

The pump passage 5 and the feeder passage 7A,
7B, a valve body (hereinafter, referred to as a seat valve body) 20 of the seat valve 300 is disposed, and the seat valve body 20 is slidably inserted into a bore 21 orthogonal to the bore 3 formed in the housing 1. It is stored. As shown in an enlarged manner in FIG. 4, the bore 21 has a bore 21 a that opens into the pump passage 5 and forms the inlet passage 22 of the seat valve 300, and a bore 21 a that has a larger diameter than the bore 21 a that opens on the outer wall surface of the housing 1. A bore portion 21b and a bore portion 2 located adjacent to the bore portion 21b.
Bore portion 21 larger in diameter than 1a and smaller than bore portion 21b
c, and an annular outlet passage 23 communicating with the feeder passages 7A and 7B is formed between the bore portions 21a and 21c. The opening end of the bore portion 21b is closed by the fixed block 2, and a hydraulic chamber 24 is formed in the bore portion 21b. 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 through 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 that connects the pump passage 5 to the feeder passages 7A and 7B via the inlet passage 22, the passages 29 and 30, and the pilot flow groove 31 (the 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.

In the fixed block 2, a bottom 4 is provided at one end.
0a (see FIG. 5), and a bore 40 having the other end opened to the outer surface of the fixed block is formed, and a spool valve body of the pilot control valve 400 (hereinafter, referred to as a pilot valve body) is slidably provided in the bore 40. 41 are arranged. As shown, the bore 40 is formed in parallel with the bore 3 of the direction switching valve 200A, and the pilot valve body 41 is also arranged in parallel with the spool valve body 4A.

As shown in the enlarged view of FIG. 5, an annular inlet passage 42 having an opening 35 and an annular outlet passage 43 having an opening 36 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. The preset force of the spring 47 is applied to the meter-in main variable throttle 16 of the directional control valve 200A as described later.
A target value of the differential pressure between A and 16B, that is, a 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.
Passages 57, 5 communicating with load passages 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, both ends of the spool valve element 4A project from the end face of the housing 1, respectively. The left end of the spool valve body 4A is connected to an operation lever (not shown) via a plug 75, and the right end of the spool valve body 4A 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 4A at the neutral position when the operation lever is not operated, and includes 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 FIG. 2, the configuration of the second and third directional control valve devices 100B and 100C is the same as that of a conventional center bypass type directional control valve. That is, the second direction switching valve device 100B has the direction switching valve 200B and the load check valve 770 incorporated in the common housing 1, and the third direction switching valve device 100C is incorporated in the common housing 1. And a load check valve 771.

The directional control valve 200B comprises a spool valve element 4B slidably inserted into a bore formed in the housing 1 as in the first embodiment.
The pump passage 772 and the load passages 773A, 7
73B, feeder passages 774A and 774B, meter-in main variable throttles 775A and 775B, meter-out main variable throttles 776A and 776B, and the like.
Downstream of the outlet side center bypass passage 751 of the directional switching valve device 100A, the inlet side center bypass passage 777 and the outlet side center bypass passage 778 connected in series to the outlet side center bypass passage 751.
And a bleed-off variable aperture (not shown). The load check valve 770 is disposed between the pump passage 772 and the feeder passages 774A and 774B.

The directional control valve 200C comprises a spool valve element 4C slidably inserted into a bore formed in the housing 1 in the same manner as described above, and a pump passage 782 and a load passage in relation to the spool valve element 4C. 783A, 783B,
Feeder passages 784A and 784B, meter-in main variable throttles 785A and 785B, meter-out main variable throttles 786A and 786B, and the like are formed. Are formed with an inlet-side center bypass passage 787 and an outlet-side center bypass passage 788 connected in series, and a bleed-off variable throttle (not shown). The load check valve 771 is disposed between the pump passage 782 and the feeder passages 784A and 784B.

The entrance side center bypass passage 750 and the exit side center bypass passage 751, the entrance side center bypass passage 777, the exit side center bypass passage 778, the entrance side center bypass passage 787, and the exit side center bypass passage 788 constitute one center bypass. Form a line,
The most downstream center bypass passage 778 is tank port 7
89 is connected to the tank 704.

The second and third directional control valve devices 10
0B and 100C have relief valves 710 for preventing the load pressure of the actuators 702 and 703 from rising above a set value.
A, 710B and 711A, 711B are incorporated.

In the hydraulic control valve device 100 of the present embodiment configured as described above, the first directional control valve device 100
A seat valve 300 is basically disclosed in Japanese Patent Application Laid-Open No. 58-501.
It operates similarly to the seat valve described in the '781 publication. That is, the pilot flow grooves 31 formed in the seat valve body 20
The opening area (opening area of the controllable restrictor 33) of the seat valve body 20 changes according to the amount of movement (stroke position) of the seat valve element 20, and the amount of movement of the seat valve element 20 depends on the pilot flow groove 31 (control variable throttle). 33) is determined according to the pilot flow rate passing through. The pilot flow rate is determined by the opening area of the variable throttle 45 of the pilot control valve 400. 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 element 20 is proportional to the pilot flow rate. Therefore, the main flow rate is determined by the opening area of the variable throttle 45 of the pilot control valve 400. Is done. Hereinafter, this will be described in more detail.

In FIG. 4, 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 Ac, the inlet pressure in the inlet chamber 22 (the supply pressure in the pump passage 5) is 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 the outer diameter of the seat portion 21a is L, 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)

In the present embodiment, 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 element 41 of the pilot control valve 400 in the valve opening direction as an urging force, and the pressure in the feeder passage 7A or 7B is applied. 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 the equation (13), Pz-PL = F (14) The main variable throttle 16A provided on the spool valve element 4A
Or, if the opening area of 16B 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.

The above equation (17) indicates that the directional control valve 200A
The flow rate (the flow rate supplied from the pump passage 5 to the load passage 6A or 6B) Qv passing through the main variable throttle 16A or 16B is determined by the supply pressure in the pump passage 5 and the load passage 6
Independent of the load pressure in A or 6B, the preset force F
And the aperture area A of the main variable stop 16A or 16B. Main variable aperture 16A at this time
Alternatively, the target value of the front-back differential pressure Pz-PL of 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 rate 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.

Further, 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 from the inlet chamber 24 to the hydraulic chamber 24 becomes 0, and the displacement of the seat valve body 20 becomes 0 from the above equation (5). 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 5 is prevented, and the seat valve 300 performs a load check function.

As described above, the hydraulic control valve device 1 of this embodiment
At 00, the first directional control valve device 200A is provided with a pressure compensation function and a load check function by the seat valve 300, whereby the following operational effects are obtained.

First, the hydraulic control valve device 1 of the present embodiment
Since the pressure compensation function is added to the first directional control valve device 200A of the first directional control device 200A, the first directional control valve device 200A
And the operability when shifting from a single operation to a composite operation is improved.

That is, in FIG. 2, the control valve device 100 of the present embodiment is used for a hydraulic circuit device of a hydraulic shovel, a hydraulic motor 701 rotates a rotary motor for rotating a rotary table, and a hydraulic actuator 702 raises and lowers a boom cylinder. Let us consider a case where a transition is made from a single swing operation for driving the swing motor 701 to a combined swing and boom raising operation for simultaneously driving the boom cylinder 702. In this case, the load pressure of the swing motor 701 in a single swing operation is relatively low, and the operation amount of the direction switching valve device 200A (the movement amount of the spool valve element 4A) is relatively small, and the swing motor 701 rotates at a very low speed. When the load pressure of the boom cylinder 702 in the combined operation is higher than the load pressure of the swing motor 701, the pump passages 5 and 772 of the directional control valve devices 200A and 200B of both actuators are connected in parallel. Therefore, more pressure oil tends to flow into the swing motor 701 which is an actuator having a low load pressure. At this time, if the first directional control valve device 10
0A is not provided with the above-described pressure compensation function, and the first directional control valve device 100A is
If the configuration is the same as that of B, an unexpected increase in the speed of the swing motor 701 is started by the composite operation. Such a behavior is caused by the second directional control valve device 200 when trying to raise the boom during crane operation while suspending and turning a fishing load.
This occurs when B is operated, and the supply flow rate to the swing motor 701 rapidly increases, and the swing speed suddenly changes, resulting in a very dangerous state.

In this embodiment, since the above-mentioned pressure compensation function is added to the first directional control valve device 100A, the flow rate Qv passing through the main variable throttle 16A or 16B is set.
Is determined by the preset force F and the opening area A of the main variable throttle 16A or 16B, regardless of the supply pressure in the pump passage 5 and the load pressure in the load passage 6A or 6B. Therefore, the sudden increase in the supply flow rate to the swing motor 701 and the sudden change in the swing speed as described above do not occur, and it is possible to safely shift from the single operation of the swing to the combined operation of the swing and the boom raising.

Second, the hydraulic control valve device 100 of this embodiment
In the first directional switching valve device 100A, two valves, a seat valve 300 and a directional switching valve 200A, are arranged in the pump passage 5, the feeder passages 7A and 7B, and the load passages 6A and 6B constituting the main circuit. Compared to a conventional hydraulic control valve device in which three valves, a closed center type directional control valve, a pressure compensating valve, and a load check valve, are arranged in the main circuit, the pressure at which the pressure oil passes through the main circuit is reduced. Loss is reduced and economical actuator operation with low energy loss is possible.

Third, in the conventional hydraulic control valve device,
In the balance piston of the pressure compensating valve, it was necessary to form many pressure receiving chambers, passages, and the like having complicated shapes. That is, it is necessary to form pressure receiving chambers at both ends of the balance piston independently of the pump passage 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, as compared with the conventional hydraulic control valve device having only the load check valve without the 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, and the valve block becomes longer. Is larger. Also, the production of the valve block becomes complicated.

In the present embodiment, the pump passage 5, the outlet passage 23 and the hydraulic chamber 24 are provided in the housing 1, and the seat valve element is located at the position where the load check valve of the conventional hydraulic control valve device without the pressure compensation function was located. 20 and housing 1
The pilot control valve 400 is arranged using the fixed block 2 that holds the seat valve body 20 separate from the pilot control valve 400. For this reason, the axial length L of the seat valve 300 at the portion where the seat valve 300 of the housing 1 is located becomes large as in a valve device having a conventional closed center type directional control valve with a pressure compensation function. Therefore, the size of the housing 1 can be reduced to the same size as that of a conventional valve device without a pressure compensation function, and the size of the housing 1 can be reduced. The fixed block 2 is also a pilot valve body 4.
1 is arranged in parallel with the spool valve element 4A to reduce the size. Therefore, the entire valve device becomes compact,
The disadvantage that it is advantageous in terms of cost and that it is difficult to mount it on a construction machine to be used is also eliminated.

Fourth, generally, the housing of such a valve device is usually made of a casting, but in the valve device of the present 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 this embodiment, the pressure compensation function is added to the center bypass type directional switching valve device, and the flow controllability is improved.
Operability when shifting from a single operation to a composite operation is improved. Further, despite the addition of the pressure compensation function, the pressure loss does not increase, and the actuator with small energy loss can be driven. 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 5 are denoted by the same reference numerals. In this embodiment, a load check function with higher liquid tightness is obtained.

In FIGS. 6 and 7, the first direction switching valve device 101A of the hydraulic control valve device 101 of the present embodiment has a seat valve 301, and the seat valve body 20 of the seat valve 301 has the structure shown in FIG. A passage 121 is formed instead of the passage 29 shown in FIG.
A check valve 122 that allows the flow of pressure oil from the inlet chamber 22 to the hydraulic chamber 24 in the passage 121 and blocks the flow in the opposite direction.
Is arranged.

In the first embodiment shown in FIGS. 1 to 5, 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 slight gap between the outer peripheral surface of b and the bore 21 to allow sliding, and when the load pressure is extremely high, a small amount of pressure oil flows from the gap into the pump passage 5 through the pilot flow groove 31. 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 the pilot passage is completely prevented. In addition, an extremely liquid-tight load check function can be obtained.

In the present 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 fixing member 2 and the housing 1.

A third 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 5 and FIGS. 6 and 7 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.

8 and 9, the first directional control valve device 102A of 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 has an adjuster screw. The adjuster screw 130 is closed at 130, and is attached to a screw hole 48 formed at the open end of the bore 40. An operation unit 131 for inserting a hexagon wrench and rotating the hex wrench is provided integrally with the head of the adjuster screw 130. Between the adjuster screw 130 and the pilot valve element 41, a spring 47 whose both ends are in contact with the pilot valve element 41 and the screw 130 is disposed, as in the first embodiment. The urging force is applied in the valve closing direction of the valve body 41.

In the present embodiment, the insertion depth of the adjuster screw 130 changes by rotating the operation section 131, 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 front-rear differential pressure of the main variable throttles 16A and 16B of the directional control valve 200A, and the main variable throttle 16
A pressure compensation characteristic of the seat valve 301 for controlling the flow rates of the passages A and 16B is set. Therefore, the target compensation differential pressure is adjusted by operating the adjuster screw 130,
The pressure compensation characteristics of the seat valve 301 can be adjusted, and the flow characteristics of the first direction switching valve device 102A can be adjusted.

Therefore, according to this embodiment, optimal pressure compensation characteristics and flow characteristics are set according to the type of actuator driven by the first direction switching valve device 102A, the type of load thereof, and the like. Operability can be improved.

A fourth embodiment of the present invention will be described with reference to FIGS.
This will be described below. In the figures, FIGS. 1 to 5 and FIGS. 6 and 7
Are given 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. .

In FIGS. 10 and 11, the first direction switching valve device 103A of 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. I have. A bore 140 having a bottom 140a at one end and an open end at the outer surface of the fixed block is formed in the fixed block 2.
A spool valve element (hereinafter, referred to as a pilot valve element) 141 of the pilot control valve 403 is slidably disposed therein. The bore 140 is also provided with the directional control valve 200 as in the previous embodiment.
A is formed parallel to the bore 3 of A (see FIG. 1). Correspondingly, the pilot valve body 141 is also arranged parallel to the spool valve body 4A (see FIG. 1).

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, a pilot variable throttle 145 that changes an opening area from a predetermined minimum opening to a predetermined maximum opening in accordance with the amount of movement of the pilot valve element 141 between the inlet passage 142 and the outlet passage 143. Has formed.

Inside the pilot valve element 141, a pressure receiving chamber 155 extending in the axial direction and opening toward the bore bottom 140a is provided.
A pressure receiving chamber 156 extending in the axial direction and opening to the pressure receiving chamber 154 side is formed, and a slidable piston 157 whose one end is in contact with the bore bottom 140 a is inserted into the open end side of the pressure receiving chamber 155. One end is a screw 1 on the open end side.
A slidable piston 158 abutting on 48 is inserted. The pilot valve element 141 has a pressure receiving chamber 155.
Radial passage 159 connecting the pressure receiving chamber 156 to the passage 151, 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 of the seat valve, the passages 36 and 37, the outlet passage 143, and the passage 1
The pressure in the feeder passage 7A or 7B is introduced through 59, and the pressure is applied in the valve closing direction of the pilot valve element 141. The passages 55 and 56 and the passage 5 are provided in the pressure receiving chamber 156.
The pressure on the high pressure side of the load passages 6A, 6B is introduced through the shuttle valve 59, the shuttle valve 59, and the passage 54, and the pressure is applied in the valve opening direction of the pilot valve body 141. Pressure receiving chamber 1
55 and 156 have the same inner diameter, and pistons 157 and 158
Also have the same outer diameter, and the pressure receiving areas of the pressure receiving chambers 155 and 156 and the pressure receiving areas of the pistons 157 and 159 are equalized.

The fixed block 2 has a passage 161 for introducing constant-pressure oil into the pressure receiving chamber 154 and a passage 162 for introducing variable-pressure oil into the passage 150. The constant pressure introduced into the pressure receiving chamber 154 is applied in the valve opening direction of the pilot valve element 141, and the pressure introduced into the pressure receiving chamber 150 is applied in the valve closing direction of the pilot valve element 141.

In the pressure receiving chamber 154, a spring 163 having one end in contact with the pilot valve body 141 and the other end in contact with the screw 146 is provided. This 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. 10, 5
Reference numeral 00 denotes a pilot pump, and a relief valve 501 is connected to a discharge pipe 507 of the pilot pump to maintain the pressure of the pilot line 502 at a constant pressure Pi. This pilot line 502 is pilot line 5
03 is connected to the above-mentioned passage 161 through a constant pressure P
i is introduced into the pressure receiving chamber 154. The pilot line 502 is connected to the primary side of the electromagnetic proportional pressure reducing valve 504,
The secondary side of the electromagnetic proportional pressure reducing valve 504 is the pilot line 50
5 is connected to the above-mentioned passage 162. 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 (FIG. 11) is
3 works as follows.

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 the equation (18), Pz-PL = Fh (19) Using the equations (12) and (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 200A is controlled by the urging force Fh regardless of the supply pressure and the load pressure.
And the opening area A of the main variable throttle 16A or 16B, 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 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 this embodiment, the urging force Fd can be adjusted by adjusting the pressure Pc, and the pressure Pc can be easily adjusted by using the controller 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 Q passing through the main variable throttle 16A or 16B of 0A
v can be controlled to further improve the operability of the actuator.

In the above embodiment, an auxiliary flow control function is added to one of the plurality of direction switching valve devices constituting the hydraulic control valve device by combining a seat valve and a pilot control valve. A similar configuration may be employed in one or all of the other directional control valve devices and an auxiliary flow control function may be added, thereby improving the flow controllability of the directional control valve device and obtaining the same effect. be able to.

[0094]

According to the present invention, a pressure compensating function and a load checking function are added to the first directional control valve device, so that a single operation of driving one of a plurality of actuators can simultaneously drive other actuators. When the operation is shifted to the combined operation, the sudden change in the supply flow rate of the pressure oil to the previous actuator can be prevented, and the operability can be improved.

Since the seat valve has a pressure compensation function and a load check function, only two valves, a seat valve and a main variable throttle, are arranged in the pump passage, the feeder passage, and the load port constituting the main circuit. Therefore, 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 becomes 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 body, it is easy to install the pilot control means, and the pilot valve body of the pilot control valve is arranged in parallel with the spool valve body. The arrangement also makes the fixed block compact.

Further, since the check valve is disposed in the pilot passage, a load check function with higher liquid tightness can be obtained.

Further, since the target compensation differential pressure indicating means is constituted by a spring, the configuration 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 a 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.

[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 a view showing opening degree characteristics of a bleed-off variable aperture, a meter-in variable aperture, and a meter-out variable aperture shown in FIG. 1;

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

FIG. 5 is an enlarged view of a pilot control valve in the hydraulic control valve device shown in FIG.

FIG. 6 is a sectional view of a hydraulic control valve device according to a second embodiment of the present invention.

FIG. 7 is a circuit diagram of a main part of the hydraulic control valve device shown in FIG.

FIG. 8 is a sectional view of a hydraulic control valve device according to a third embodiment of the present invention.

9 is a circuit diagram of a main part of the hydraulic control valve device shown in FIG.

FIG. 10 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.

11 is a circuit diagram of a main part of the hydraulic control valve device shown in FIG.

[Explanation of symbols]

REFERENCE SIGNS LIST 100 Hydraulic control valve device 100A to 100C First to third directional switching valve devices 200A Directional switching valve 300 Seat valve 400 Pilot control valve 1 Housing 2 Fixed block 3 Bore 4A Spool valve element 5 Pump passage 6A, 6B Load port 7A, 7B Feeder passage 16A, 16B Main variable throttle 20 Seat valve body 21 Bore 31 Pilot flow groove 33 Control variable throttle 35, 36, 37 Pilot passage 40 Bore 41 Pilot valve body 45 Pilot variable throttle 47 Spring (target compensation differential pressure indicating means) 50, 51 Pressure receiving chamber (biasing means) 754A, 754B Variable throttle for bleed-off 101 Hydraulic control valve device 101A First directional switching valve device 301 Seat valve 122 Check valve 102 Hydraulic control valve device 102A First directional switching valve Device 401 Pilot control valve 30 adjuster screw 131 operating section 103 hydraulic control valve apparatus 103A first directional control valve device 403 pilot control valve 150, 154 receiving chamber (target compensation differential pressure command unit; hydraulic force generating means)

──────────────────────────────────────────────────続 き Continuation of the front page (72) Inventor Masami Ochiai 650, Kandachicho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura Plant (56) References JP-A-4-8902 (JP, A) JP-A-53 -41681 (JP, A) JP-A-2-248701 (JP, A) (58) Fields investigated (Int. Cl. ⁷ , DB name) F15B 11/00-11/22

Claims (9)

(57) [Claims] A housing; a pump port formed in the housing and connected to a hydraulic pressure source and a tank port connected to a tank; and a first spool valve body slidably disposed in the housing. A pump passage formed in the housing and communicating with the pump port, a feeder passage communicating with the pump passage, a load passage having a load port connected to the first actuator, and a discharge passage communicating with the tank port; A first center bypass passage formed in the housing and communicating the pump port with the tank port; and a movement amount of the first spool valve body positioned between the feeder passage and the load passage. And a main variable throttle that changes an opening area according to the first spool valve body located on the center bypass passage. A first direction switching valve device of a center bypass type having a bleed-off main variable throttle for changing an opening area according to a moving amount; a second spool valve body slidably disposed in the housing; A second center bypass passage connected in series with the first center bypass passage and controlling a flow rate of pressure oil supplied from the pump port to a second actuator. A first directional control valve device, wherein the first directional control valve device is connected to the main variable throttle via the feeder passage from the pump passage according to a pressure difference between the front and rear of the main variable throttle. Auxiliary flow rate control means for auxiliary control of the flow rate of the pressure oil supplied to the pump, wherein the auxiliary flow rate control means comprises: (a) the auxiliary flow rate control means between the pump passage and the feeder passage; A seat valve body slidably disposed in a 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 a controllable throttle in the pump passage. A seat passage which has a pilot passage communicating with the feeder passage through the pilot passage, and determines 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, And a pilot control means for controlling the pilot flow rate in accordance with the differential pressure across the variable throttle. 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 first direction switching valve device is disposed in the pilot passage, and includes a check valve for preventing a backflow of the pressure oil in the pilot passage. A hydraulic control valve device further provided. 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.