JP2987279B2 - Hydraulic control valve device and hydraulic drive 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 having a spool type flow control valve and having an auxiliary flow control function and a load check function. The present invention relates to a device and a hydraulic drive device incorporating the hydraulic control valve device.

[0002]

2. Description of the Related Art The following prior arts are known as hydraulic control valve devices or hydraulic drive devices used for construction machines such as hydraulic excavators.

JP-A-60-5928 JP-A-62-38496 JP-A-59-51861 JP-A-60-11706 JP-A-2-134402 JP-A-58-501781 JP-A-60-5928, JP-B-62-384
No. 96 and Japanese Utility Model Laid-Open Publication No. 59-51861 describe a hydraulic drive device using a valve device having a center bypass type flow control valve. The center bypass type flow control valve is such that the center bypass passage connecting the pump port to the tank is throttled according to the amount of movement of the spool, and the discharge pressure of the pump rises by narrowing the center bypass passage. The pressure oil is supplied to the actuator through the feeder passage and the meter-in variable throttle. A load check valve is installed in the feeder passage to prevent backflow of pressure oil.

[0004] Also, Japanese Patent Application Laid-Open No. 60-5928,
JP-B-62-38496, JP-A-59-518
No. 61 discloses a hydraulic drive device for a hydraulic excavator. Japanese Patent Application Laid-Open No. 60-5928 discloses a hydraulic drive device in which an oil flow control valve for supplying pressure oil to an arm cylinder of a hydraulic excavator and a pump port. Is connected to a swing priority switching valve that operates in accordance with the swing pilot pressure, and supplies pressure oil to the swing motor. When the swing flow control valve and the arm flow control valve are operated simultaneously, supply them to the arm flow control valve. The pressure of the pressurized oil supplied to the swirling flow control valve is increased by reducing the flow rate of the pressurized oil to be supplied.

In the hydraulic drive system disclosed in Japanese Patent Publication No. 62-38496, for the same purpose, a variable sequence valve which operates according to the turning pilot pressure is connected between the arm flow control valve and the pump port. ing.

In the hydraulic drive described in Japanese Utility Model Laid-Open Publication No. Sho 59-51861, for the same purpose, the load check valve installed at the inlet port of the arm flow control valve is connected to the pressure at the inlet port of the swirling flow control valve. As the pilot pressure, and the poppet of the load check valve is pushed and moved by the pilot pressure to reduce the flow rate through the load check valve.

Japanese Unexamined Patent Publications Nos. 60-11706 and 2-134402 disclose a hydraulic drive device using a valve device having a closed center type flow control valve. The closed center type flow control valve is designed to prevent the pump port from communicating with the tank regardless of the position of the spool, and is usually used in combination with a load sensing system that controls the discharge flow rate of the hydraulic pump according to the load pressure. Used.
In addition, a pressure compensating valve is provided upstream of the closed center type flow control valve so that the operating speed of the plurality of actuators does not change due to a change in load. A load check valve for preventing backflow of pressure oil is disposed between the pressure compensating valve and the flow control valve.

In the valve device described in Japanese Patent Application Laid-Open No. 2-134402, when a hydraulic control valve device is constructed by combining a spool type flow control valve, a pressure compensating valve and a load check valve, the number of pipes is reduced. For the purpose of reducing the size and reducing the size, these three valves are incorporated in one block to constitute one valve device.

On the other hand, Japanese Patent Laying-Open No. 58-501781 proposes a seat valve type hydraulic control valve device instead of a spool type. This hydraulic control valve device is composed of a combination of a seat valve and a pilot control valve.

[0010]

Problems to be Solved by the Invention
No. 8 and JP-A-62-38496, a load check valve and a swing priority switching valve or a variable sequence valve are installed in a main flow path upstream of an arm flow control valve, JP-A-60-117
Also in the hydraulic control valve device described in Japanese Patent Application Laid-Open No. 2006-134402 and Japanese Patent Application Laid-Open No. 2-134402, a load check valve and a pressure compensating valve are provided in the main flow path upstream of the flow control valve. The swing priority switching valve, the variable sequence valve, and the pressure compensating valve each perform one type of auxiliary flow control function for the arm flow control valve. However, due to the additional installation of these valves, the hydraulic oil supplied to the actuator from the hydraulic pump passes through these valves, the load check valve, and the flow control valve (main variable throttle). There is a problem that the pressure loss increases due to the flow resistance of the valve, and the energy loss increases.

In the valve device described in Japanese Utility Model Laid-Open No. 59-51861, the flow rate is reduced by the load check valve, so that no special valve is additionally provided, and the pressure loss is smaller than that of the above-described valve device. However, since the poppet of the load check valve is only pushed and moved by the pilot pressure, it is impossible to accurately control the flow rate of the load check valve, and it is not possible to obtain an auxiliary flow rate control function with high control accuracy.

Further, in the valve device described in Japanese Patent Application Laid-Open No. 2-134402, 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.

The hydraulic control valve device described in Japanese Patent Application Laid-Open No. 58-501781 uses a seat valve type in place of the spool type flow control valve. Type of flow control valve cannot be used.

In the hydraulic control valve device described in Japanese Patent Application Laid-Open Nos. 60-5928 and 62-38496, the swing pressure between the arm flow control valve and the pump port depends on the pilot pressure. By providing a turning priority changeover valve or a variable sequence valve that operates by turning, the pressure of the pressure oil supplied to the turning direction changeover valve is increased, and the operability in the combined operation of the arm and the turning is improved. However, in this prior art, since the swirling priority switching valve or the variable sequence valve is disposed on the pump line common to the other flow control valves, the influence of the operation of the rotation priority switching valve or the variable sequence valve affects the flow control valve for the arm. When the simultaneous operation of the swirling flow control valve and the other flow control valve is performed, the operation of the swirling priority switching valve or the variable sequence valve impairs the complex operability.

A first object of the present invention is to provide a hydraulic control valve device having a spool-type flow control valve, which has an auxiliary flow control function with high control accuracy, increases pressure loss and enlarges the structure. The object of the present invention is to provide a hydraulic control valve device and a hydraulic drive device without any.

A second object of the present invention is to provide a hydraulic control valve device having a center bypass type flow control valve,
An object of the present invention is to provide a hydraulic control valve device and a hydraulic drive device that can supplementarily control a supply flow rate to only a target flow control valve in a combined operation in which a plurality of actuators are simultaneously driven and can improve combined operability.

To achieve the first and second objects, according to a first aspect of the present invention, a housing, a pump passage formed in the housing, and at least one of a pump passage incorporated in the housing. Directional switching valve means,
The direction switching valve means is slidably disposed in the housing to form a pair of main variable throttles and is formed in the housing, and a main spool constituting a flow control valve,
A feeder passage for supplying pressure oil from the pump passage to the pair of main variable throttles, and a pair of load passages formed in the housing and through which the pressure oil that has passed through the pair of main variable throttles respectively flows. Wherein the direction switching valve means restricts a flow rate of the pressure oil supplied from the pump passage to the pair of main variable throttles via the feeder passage, and the pair of load passages. Further comprising auxiliary flow rate control means for auxiliaryly controlling the flow rate of the pressure oil flowing into the housing, wherein the auxiliary flow rate control means is: (a) a seat valve disposed in the feeder passage, which moves into the housing. is freely positioned, and the seat valve body forming an auxiliary variable throttle in said feeder passage, formed in the seat valve body, it possesses a diaphragm control variable to vary the opening area in accordance with the amount of movement of said seat valve body ,And Moved the serial main spool from the neutral position
When the pair of main variable throttles is open,
The upstream side of the auxiliary variable throttle via the control variable throttle contact downstream of the feeder passage, through its (b) the feeder passage; seat valve and which have a load check function to prevent backflow of A pilot line for determining an amount of movement of the seat valve body according to a flow rate of the pressure oil;
(C) having a pilot spool having a pilot variable throttle disposed on the pilot line and an input means for inputting a flow restriction signal,
Serial wherein by moving the pilot spool pilot <br/> changing the opening area of the bets variable throttle, and the pilot spool valve <br/> for controlling the flow rate of the hydraulic fluid flowing through said pilot line; wherein the sheet The valve is orthogonal to the main spool.
The pilot spool valve
The spool is arranged parallel to the main spool.
Hydraulic control valve and wherein the has is provided.

In the above hydraulic control valve device, preferably, the direction switching valve means further includes a fixed block for holding the seat valve body in the housing via a spring, and the pilot spool of the pilot spool valve is It is incorporated in the fixed block.

[0019]

[0020] Preferably, the feeder passage is provided with:
A first passage portion located upstream of the auxiliary variable throttle and communicating with the pump passage; and a pair of the first passage portion located on both sides of the first passage portion downstream of the auxiliary variable throttle of the feeder passage. And a second and a third passage portion communicating with the main variable throttle. The seat valve is disposed at a connection point between the first passage portion and the second and the third passage portions. Further, preferably, the opening degree characteristic is set so that the control variable throttle is slightly opened at the fully closed position of the seat valve, and the direction switching valve means is disposed in the pilot line to prevent a backflow of the pressure oil. Further comprising a check valve,
The check valve is incorporated in the seat valve body.

Preferably, the hydraulic control valve device further comprises a plurality of spool type direction switching valve means incorporated in the housing, at least one of which has the auxiliary flow rate control means. Means.

The input means of the pilot spool valve has, for example, a passage for inputting a pressure signal generated outside the direction switching valve means as the flow rate limiting signal.

In the hydraulic control valve device, preferably, the pilot spool valve includes a pilot spool that forms the pilot variable throttle, and a first urging force that applies a predetermined urging force to the pilot spool in a valve opening direction.
And a second biasing means connected to the input means for applying a biasing force to the pilot spool in accordance with the flow rate restriction signal in a valve closing direction.

Preferably, the first biasing means has a spring which biases the pilot spool in a valve opening direction with a predetermined preset force. In this case, preferably, the path
The pilot spool valve further includes operating means for externally adjusting the preset force of the spring.

Preferably, the first biasing means has at least one pressure receiving chamber for applying a predetermined hydraulic pressure to the pilot spool in a valve opening direction.

Still preferably, the second biasing means has at least one pressure receiving chamber for applying a hydraulic pressure in the valve closing direction based on the flow rate restriction signal to the pilot spool.

In the above-mentioned hydraulic control valve device, preferably, the input means includes a passage for introducing a signal generated outside the direction switching valve means as the flow rate limiting signal to the second urging means. Have.

In this case, preferably, the first urging means has a pressure receiving chamber into which the inlet pressure of the pair of main variable throttles is introduced.

Preferably, the first urging means has a pressure receiving chamber into which the pressure of the pump passage is introduced.

According to a second aspect of the present invention, there is provided a hydraulic pump, comprising: a plurality of hydraulic actuators driven by hydraulic oil discharged from the hydraulic pump; And at least first and second direction switching valve means each having a spool type flow control valve which is operated in accordance with an operation signal and controls the flow rate of the hydraulic oil supplied to each of the plurality of hydraulic actuators. A hydraulic control valve device according to the first concept, wherein at least the first direction switching valve means is a direction switching valve means having the auxiliary flow rate control means; and
Of externally generated directional control valve means, said this pilot
And a signal generating and transmitting means to be introduced into the input means of the tospool valve .

In the above hydraulic drive device, preferably,
The signal generation and transmission means includes means for detecting an operation signal given to the second direction switching valve means, and means for introducing the operation signal as the flow rate restriction signal to the input means of the pilot spool valve .

Preferably, the signal generating and transmitting means is a setting means operated by an operator to output a setting signal, a means for generating a control signal corresponding to the setting signal, and the control signal is transmitted to the flow rate limiting signal. As the pyro
Means to be introduced into the input means of the throttle valve .

More preferably, the signal generation / transmission means is a means for outputting a setting signal operated by an operator, and a control signal corresponding to the operation signal given to the second direction switching valve means and the setting signal. Means for generating the pilot signal and the pilot signal as the flow rate limiting signal .
Means for introducing to the input means of the pool valve .

In the above-mentioned hydraulic drive device, preferably,
The flow control valve is a center bypass type spool valve.

[0035]

The operation of the hydraulic control valve device and the hydraulic drive device of the present invention configured as described above is as follows.

In the hydraulic control valve device of the present invention, when the spool valve is moved from the neutral position, one of the main variable throttles opens,
Most of the pressure oil upstream of the feeder passage passes through the seat valve as the main flow rate and flows out downstream of the feeder passage, and the remaining pressure oil upstream of the feeder passage passes through the pilot line as the pilot flow rate. The oil flows out downstream of the feeder passage, merges with the main flow rate, and the merged pressure oil passes through the main variable throttle and is supplied to the load port. On the other hand, the seat valve operates according to the principle described in Japanese Patent Application Laid-Open No. 58-501781, and the amount of movement of the seat valve body is determined according to the pilot flow rate passing through the controllable throttle. This pilot flow is controlled by the pilot spool valve in accordance with the flow rate limit signal. That is, the moving amount of the seat valve element is determined according to the flow rate limiting signal, and the main flow rate passing through the seat valve is adjusted. In this way, the flow rate of the pressure oil supplied to the main variable throttle is limited, and the flow rate of the pressure oil flowing into the load passage is accurately and auxiliary controlled.

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, the seat valve body is urged in the valve closing direction, and the seat valve is urged. Is fully closed. Therefore, the backflow of the pressure oil is prevented, and the load check function is performed.

As described above, in the hydraulic control valve device of the present invention,
By arranging a seat valve in a feeder passage where a load check valve has conventionally been provided, two functions of an auxiliary flow control and a load check are performed. For this reason, the valve device of the present invention has an auxiliary flow control function and requires only the same pressure loss as the conventional hydraulic control valve device without the auxiliary flow control function, and increases the pressure loss by providing the auxiliary flow control function. Is avoided.

Further, in the hydraulic control valve device of the present invention, the seat valve which performs the two functions of the auxiliary flow control and the load check as described above is disposed in the feeder passage where the conventional load check valve is provided, and the pilot spool valve is provided. Can be installed in parts other than the housing. For this reason, the configuration around the seat valve body of the seat valve is simplified, and the axial length (the size in the direction perpendicular to the spool valve) of the seat valve body at the portion where the seat valve body is located in the housing becomes large. This makes the housing compact and easy to manufacture. In the hydraulic control valve device of the present invention,
The pilot spool of the pilot spool valve is the main spool.
The valve system is arranged in parallel with the
Become a pact.

As described above, in the hydraulic control valve device provided with the spool-type flow control valve, the auxiliary flow control function with high control accuracy was achieved without increasing the pressure loss or increasing the size of the structure. The first object is achieved.

The flow control valve constituted by the main spool is of a center bypass type, and an external signal is introduced as the flow control signal, so that a target flow control valve is provided in a combined operation for simultaneously driving a plurality of actuators. The second object of the present invention is achieved by supplementarily controlling the supply flow rate to only the second supply.

Further, the flow control valve constituted by the main spool is a center bypass type, and a differential pressure across the main variable throttle is introduced as the flow rate limiting signal, thereby providing a center bypass type flow control valve. The device is provided with a pressure compensation function.

The flow control valve constituted by the main spool is a closed center type, and a differential pressure before and after the main variable throttle is introduced as the flow restriction signal, thereby providing a closed center type flow control valve. The device is provided with a pressure compensation function without increasing the pressure loss or increasing the size of the structure.

[0044] The seat valve body via a spring which fixed block held in the housing, pies to the fixed block
By incorporating a lot spool valve, the pilot spur can be
A valve can be installed, and the housing becomes compact as described above. At this time, by arranging the spool of the pilot spool valve in parallel with the main spool,
The fixed block itself is also compact.

By setting the opening degree characteristic so that the controllable throttle is slightly opened at the fully closed position of the seat valve, the generation of the pilot flow rate is stabilized and the processing of the controllable throttle becomes easy. At this time, by installing a check valve in the pilot line, a highly liquid-tight load check function can be obtained. Since the check valve is arranged in the pilot line, the pressure loss in the main circuit (feeder passage) does not increase.

When the first biasing means of the pilot spool valve is constituted by a spring, when the pressure compensation control is performed as the auxiliary flow control, the target compensation differential pressure is set by the preset force of the spring. By making the preset force of the spring adjustable from the outside, the desired compensation differential pressure can be arbitrarily adjusted. When the first urging means is constituted by a pressure receiving chamber, the target compensation differential pressure is set by the hydraulic pressure applied by the pressure receiving chamber. Also in this case, the target compensation differential pressure can be adjusted by introducing a variable pressure into the pressure receiving chamber, and the pressure can be easily adjusted by using an electromagnetic proportional pressure reducing valve or the like. It becomes possible.

By introducing an external oil pressure signal as a flow rate limiting signal to the second urging means, the action relating to the second object is achieved. That is, the supply flow rate to only the intended flow rate control valve can be supplementarily controlled.

In this case, in the second concept of the present invention,
By introducing an operation signal given to the second direction switching valve means as an external signal (flow rate limiting signal) to the second biasing means, the second direction switching valve is provided in a combined operation for simultaneously driving a plurality of actuators. The passing flow rate of the first direction switching valve means can be automatically controlled in accordance with the magnitude of the operation signal of the means, and the composite operability is improved.

In the second concept of the present invention, the flow rate limiting signal is generated based on the setting signal from the setting means operated by the operator, so that the flow rate of the supply to the actuator can be assisted by the operator. It can be controlled, and the composite operability is further improved.

Further, according to the second concept of the present invention, the setting signal from the setting means operated by the operator and the second signal
The flow rate limiting signal is generated based on the operation signal given to the direction switching valve means of the second direction switching valve means, and the supply flow rate to the actuator is supplementarily controlled according to the intention of the operator and the operation signal of the second direction switching valve device. It can be controlled, and the composite operability is further improved.

At this time, by introducing the inlet pressure of the pair of main variable throttles into the pressure receiving chamber of the first urging means, the pilot spool operates only when the load pressure of the actuator is low, and the pilot spool operates. Reduces the flow rate and selectively exerts the auxiliary flow control function. This avoids unnecessary energy loss while ensuring good composite operation.

By introducing the pump pressure into the pressure receiving chamber of the first biasing means, when the pump pressure increases to a predetermined pressure determined by the relationship with the flow rate limiting signal, the control amount of the pilot flow rate decreases, and the second control rate decreases. The pressure of the pressure oil supplied to the direction switching valve means (the driving pressure of the corresponding actuator) changes according to the flow rate restriction signal, thereby further improving the combined operability.

[0053]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Some embodiments of the present invention will be described below with reference to the drawings. In these embodiments, the first to ninth embodiments apply the present invention to a valve device having a center bypass type flow control valve, and the tenth to fourteenth embodiments use a closed center type flow rate control valve. The present invention is applied to a valve device provided with a control valve. In addition, the first to fifth
Is an embodiment in which an external signal is used as a flow restriction signal provided to the pilot flow control means. In the sixth to fourteenth embodiments, a differential pressure across the main variable throttle is used as a flow restriction signal, and pressure compensation control is performed. Is what you do.

First Embodiment First, a first embodiment of the present invention will be described with reference to FIGS. The present embodiment to the fifth embodiment use an external signal as the flow rate limiting signal in the valve device having the center bypass type flow control valve as described above.

In FIGS. 1 and 2, the hydraulic control valve device of the present embodiment is generally designated by the reference numeral 100, and this hydraulic control valve device 100 is supplied to a hydraulic actuator 701 as shown in FIG. Directional switching valve device 100A for controlling the flow of pressurized oil, hydraulic actuator 7
02 has a second direction switching valve device 100B for controlling the flow of pressure oil supplied to the hydraulic actuator 703, and a third direction switching valve device 100C for controlling the flow of pressure oil supplied to the hydraulic actuator 703. ing.

The hydraulic control valve device 100 includes a housing 1 common to the first to third directional switching valve devices, and a first directional switching valve device 1 integrally mounted on the housing 1.
The first directional control valve device 100A includes a main spool valve 2 which is incorporated in the housing 1 and constitutes a center bypass type flow control valve.
00A and the seat valve 30 incorporated in the housing 1
0, and a pilot spool valve 400 incorporated in the fixed block 2 to constitute a pilot flow control valve.

The main spool valve 200A, the seat valve 300, and the pilot spool valve 400 are configured as follows.

A bore 3 is formed through the housing 1 and a main spool 4A of the main spool valve 200A is formed in the bore 3.
Are slidably inserted. In the housing 1, a pump passage 5 having a pump port 5a (see FIG. 2) connected to a hydraulic pump 700, and load ports 6a and 6b connected to a hydraulic actuator 702 (see FIG. 2).
Are formed, and a feeder passage 7 which is branched from the pump passage 5 and can communicate with the load passages 6A, 6B is formed. The feeder passage 7 is a passage portion 7C communicating with the pump passage 5, and a passage portion 7C. A pair of passage portions 7A and 7B located on both sides of the passage portion 7C; a passage portion 7C and a passage portion 7;
A, 7B, and an annular passage portion 23 that communicates therewith. Hereinafter, the passage portions 7A to 7C and 23 are each simply referred to as a feeder passage.

Near the center of the bore 3, 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.
In the bore 3, annular feeder passages 8A and 8B forming part of the feeder passages 7A and 7B, annular load passages 9A and 9B forming part of the load passages 6A and 6B, and a tank port 8 are provided.
5 (see FIG. 2) in the form of annular discharge passages 10A, 10A.
B are formed, 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, and the feeder passage 8B
Between the load passage 9B and the load passage 9B and the discharge passage 10
Lands 11B and 12B are formed between the lands 11B and 12B. The tank port 85 is connected to the tank 704.

Notches 753A, 75 are provided on the main spool 4A.
3B and a cylindrical portion 755 are formed. Notch 753
A and the cylindrical portion 755 are the land portions 752A and 752B.
In cooperation therewith, a bleed-off variable throttle 754A located between the entrance-side center bypass passage 750 and the exit-side center bypass passages 751A and 751B is formed, and this variable throttle 754A is shown in PT in FIG. As described above, the opening area is changed from the fully open position to the fully closed position according to the moving amount (spool stroke) of the main spool 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 the entrance center bypass passage 75
0, and a bleed-off variable throttle 754B located between the outlet side center bypass passages 751B and 751A. The variable throttle 754B is located on the left side of the main spool 4A as shown in PT of FIG. The opening area is changed from the fully open position to the fully closed position according to the amount of movement of.

The main spool 4A has notches 14A,
14B and notches 15A and 15B are formed. The notch 14A cooperates with the land 11A to form a meter-in main variable throttle 16A located between the feeder passage 8A and the load passage 9A.
As shown at P-A, the opening area is changed from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the main spool 4A to the right in the drawing. The notch 14B cooperates with the land portion 11B to form a meter-in main variable throttle 16B located between the feeder passage 8B and the load passage 9B. The variable throttle 16B is as shown in P-A of FIG. And the main spool 4
The opening area is changed from the fully closed position to a predetermined maximum opening according to the amount of movement of A in the left of the drawing. The notch 15B cooperates with the land portion 12B to form a meter-out main variable throttle 17 located between the load passage 9B and the discharge passage 10B.
B, the variable throttle 17B changes the opening area from the fully closed position to a predetermined maximum opening according to the amount of movement of the main spool 4A to the left as shown in BT of FIG. . The notch 15A forms a meter-out main variable throttle 17A located between the load passage 9A and the discharge passage 10A in cooperation with the land 12A.
Changes the opening area from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the main spool 4A to the right as shown in BT of FIG.

The feeder passage 7C and the feeder passage 7
A valve body (hereinafter, referred to as a seat valve body) 20 of the seat valve 300 is disposed in a portion of the annular feeder passage 23 which is a connection point between the seat valves A and 7B, and the seat valve body 20 is a bore formed in the housing 1. It is slidably housed in a bore 21 orthogonal to 3. As shown in an enlarged manner in FIG. 4, the bore 21 has a bore portion 21a serving also as a part of the feeder passage 7C.
And a bore portion 21b opened to the outer wall surface of the housing 1 and having a larger diameter than the bore portion 21a, and a bore portion 21c located adjacent to the bore portion 21b and having a diameter larger than the bore portion 21a and smaller than the bore portion 21b. And the bore portions 21a, 21c
The above-mentioned annular feeder passage 23 is located therebetween. The opening end of the bore 21b is closed by the fixed block 2, and a hydraulic chamber 24 is formed in the bore 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 includes a seat portion 20a that can touch the edge portion between the bore portion 21a and the annular feeder passage 23, and a sliding portion 20c located in the bore portion 21a below the seat portion 20a. , A sliding portion 20b located in the bore portions 21b and 21c, and the sliding portion 20b corresponding to the above-mentioned diameter relationship between the bore portion 21a and the bore portion 21c.
Has a larger diameter than the sliding portion 20c. Sliding part 20c
Is formed in a cylindrical shape having a recess 26 in the center as shown in the figure, and a plurality of semicircular notches 27 are formed through the cylindrical side wall, and the notches 27 cooperate with the seat portion of the housing 1. This forms an auxiliary variable throttle 28 located between the feeder passage 7C and the feeder passage 23. The auxiliary variable throttle 28 changes the opening area from the fully closed position to a predetermined maximum opening according to the movement amount (stroke) of the seat valve element 20, as indicated by FF in FIG.

A pilot flow groove 31 communicating with the feeder passage 7C through passages 29 and 30 formed inside the seat valve body 20 is formed on the outer peripheral surface of the sliding portion 20b of the seat valve body 20. . The pilot flow groove 31 forms a controllable throttle 33 located between the feeder passage 7C and the hydraulic chamber 24 in cooperation with a land 32 formed by a step between the bore 21c and the bore 21b. . The control variable throttle 33 is formed at a position where the seat valve element 20 is completely closed by the land portion 23 when the seat valve element 20 is at the fully closed position. As shown by F-C in FIG. The opening area is changed from the fully closed position shown to a predetermined maximum opening degree according to (stroke).

Returning to FIG. 1, the fixed block 2 has a hydraulic chamber 2
A passage 35 communicating with the feeder passage 4 and a passage 36 communicating with the feeder passage 23 via a passage 37 formed in the housing 1 are formed, and a pilot spool valve 400 is disposed between the passage 35 and the passage 36. . The passages 35 to 37, the hydraulic chamber 24, the passages 29, 30 and the pilot flow groove 31 connect the feeder passage 7C to the feeder passages 23, 7A, 7B via the controllable throttle 33, and the pressure oil flowing therethrough A pilot line for determining a moving amount, that is, a stroke of the seat valve body 20 based on the flow rate is formed.

The fixed block 2 has a bottom 40a at one end.
A bore 40 having the other end (see FIG. 6) and having an open end on the outer surface of the fixed block is formed, and the spool 41 of the pilot spool valve 400 is slidably disposed in the bore 40. As shown, the bore 40 is formed in parallel with the bore 3 of the main spool valve 200A, and the pilot spool 41 is also arranged in parallel with the main spool 4A.

As shown in the enlarged view of FIG. 6, 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. An annular land portion 44 is located between the air passage 43 and the outlet passage 43. Inlet passage 42 and outlet passage 4
3 also constitutes a part of the pilot line. The pilot spool 41 includes a spool portion 41 a located on the bore bottom portion 40 a side, a spool portion 41 b located on the opening end side of the bore 40, and a small diameter portion 4 located near the land portion 44.
1c and an inclined portion 41d connecting the small diameter portion 41c and the spool portion 41a. The inclined portion 41d cooperates with the land portion 44 to form a pilot variable throttle 45 located between the inlet passage 42 and the outlet passage 43. As shown in FIG.
The opening area is changed from a predetermined minimum opening to a predetermined maximum opening in accordance with the movement amount (stroke) of 1.

The open end of the bore 40 is
And the screw 46 and the pilot spool 41 are closed.
A spring 47 is disposed between both ends of which abuts against the pilot spool 41 and the screw 46 to urge the pilot spool 41 in the valve closing direction. Screw 46
Is attached to a screw hole 48 formed in the opening end portion of the bore 40, and a preset force is applied to the spring 47 by the screw 46.

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. Passages 80 opening to the pressure receiving chambers 50 and 51 respectively are provided in the fixed block 2.
0,801 are formed. The passage 800 is connected via a line 803 to a shuttle valve 802 for taking out a pilot pressure P2a or P2b, which is an operation signal of the main spool valve 200B of the second direction switching valve device 100B, so that the pilot pressure is applied to the pressure receiving chamber 50. P2a or P2b is introduced, and is applied in the valve closing direction of the pilot spool 41. The passage 801 is connected to the tank 704 via the line 804, and maintains the pressure receiving chamber 51 at the tank pressure. Accordingly, pilot spool valve 400 controls the pilot flow rate flowing through the pilot line according to pilot pressure P2a or P2b of main spool valve 200B.

Returning to FIG. 1, both ends of the main spool 4A project from the end face of the housing 1, respectively. The left end of the main spool 4A in the drawing is located in a pressure receiving chamber 811 formed by a cover 810 attached to the housing 1, and the cover 810 has a main spool valve 20 connected to the pressure receiving chamber 811.
A passage 812 for introducing a pilot pressure P1a, which is an operation signal of 0A, is formed. The right side end of the main spool 4A is connected to a centering spring mechanism 77 via a plug 76. As is well known, the centering spring mechanism 77 holds the main spool 4A at the neutral position when the operation lever is not operated.
One spring 78 and two seats 79 and 80 are formed. The centering spring mechanism 77 is located in a pressure receiving chamber 813 formed by the cover 81 attached to the housing 1, and the cover 81 has a passage 814 for introducing pilot pressure P1b, which is an operation signal of the main spool valve 200A, into the pressure receiving chamber 813. Is formed. The main spool 4A moves rightward in the figure when the pilot pressure P1a is introduced into the pressure receiving chamber 811.
Is moved to the left in the figure by being introduced into the pressure receiving chamber 813.

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 flow control valve. That is, the second direction switching valve device 100B has the main spool valve 200B and the load check valve 770 which are incorporated in the common housing 1 and constitutes a flow control valve, and the third direction switching valve device 100C is common. It has a main spool valve 200C and a load check valve 771 which are incorporated in the housing 1 and constitute a flow control valve.

The main spool valve 200B has a main spool 4B slidably inserted into a bore formed in the housing 1 as in the first embodiment, and the load passages 773A and 773A are connected to the main spool 4B. 773B, feeder passage 774
(774A, 774B, 774C), meter-in main variable throttles 775A, 775B, meter-out main variable throttles 776A, 776B, etc. are formed, and the pump port 5a and the inlet side center bypass of the first direction switching valve device 100A are formed. An inlet-side center bypass passage 777 and an outlet-side center bypass passage 778 connected in series to the passage 750, respectively, and a bleed-off variable throttle (not shown) are formed. The feeder passage 774 branches off from the pump passage 5, and the load check valve 770 is connected to the feeder passage 77.
4C and the feeder passages 774A and 774B.

The main spool valve 200C has a main spool 4C slidably inserted into a bore formed in the housing 1 in the same manner as described above, and the load passages 783A and 783B and the feeder passage are associated with the main spool 4C. 784 (784
A, 784B, 784C), meter-in main variable aperture 7
85A, 785B and meter-out main variable aperture 786
A, 786B, etc. are formed, and the downstream center bypass passage 751 of the first direction switching valve device 100A is downstream of the upstream center bypass passage 751 and is connected in series with the inlet side center bypass passage 78.
7 and an outlet side center bypass passage 788 and a bleed-off variable throttle (not shown) are formed. The feeder passage 784 is branched from the pump passage 5, and the load check valve 771 is connected to the feeder passage 784C and the feeder passage 784.
A, 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 have a single center bypass. Form a line,
The most downstream center bypass passage 788 is tank port 8
5 is connected to the tank 704.

In FIG. 2, the first and third directional control valve devices 100A and 100C are provided with an actuator 70.
A relief valve 710A, 710B and 711A, 711B for preventing the load pressure of 2,703 from rising above a set value is incorporated. The relief valves 710A and 710B of the first direction switching valve device 100A are not shown in FIG.

In the hydraulic control valve device 100 of the present embodiment configured as described above, the first directional control valve device 100
The seat valve 300 of A operates on the principle described in JP-A-58-501781. That is, the seat valve element 20
The opening area of the pilot flow groove 31 formed on the land portion 32 (the opening area of the controllable throttle 33) changes according to the movement amount (stroke) of the seat valve body 20, and the movement amount of the seat valve body 20 It is determined according to the pilot flow rate passing through the flow groove 31 (the controllable throttle 33). Also, the pilot flow rate is
It is determined by the opening area of the variable stop 45 of 0. As a result, the main flow flowing from the feeder passage 7C to the feeder passage 23 via the auxiliary variable throttle 28 of the seat valve body 20 is proportional to the pilot flow, and the main flow is the opening area of the variable throttle 45 of the pilot spool valve 400. Is determined.

In the pilot spool valve 400, the opening area of the variable throttle 45 is controlled so that the pilot pressure P2a or P2b, which is an external signal, is changed in accordance with the flow rate limiting signal.

As described above, the seat valve 300 is supplied from the pump passage 5 to the main variable throttle 16A or 16B via the feeder passage 7 in combination with the pilot lines 24, 29 to 31, 35 to 37 and the pilot spool valve 400. The flow rate of the pressurized oil to be controlled
2b (flow rate limiting signal), and performs an auxiliary flow rate control function of auxiliary controlling the flow rate of the pressure oil flowing into the pair of load passages 6A, 6B. Hereinafter, this will be described in more detail.

First, in FIG. 4, the effective pressure receiving area of the end face of the sliding portion 20c located in the feeder passage 7C of the seat valve element 20 is Ap, and the effective pressure receiving area of the annular portion located in the annular feeder passage 23 is Az. The effective pressure receiving area of the end face of the sliding portion 20b located in the hydraulic chamber 24 is Ac, the pressure of the feeder passage 7C (the supply pressure in the pump passage 5) is Pp, the pressure in the feeder passage 23 is Pz, and the hydraulic chamber 24 is Pz. Assuming that the internal pressure is Pc, the pressure receiving areas Ap, Az,
From the balance of Ac, Ac = Az + Ap (1) is established, and from the balance of the pressure applied to the seat valve body 20, Ap / Pp + Az · Pz = Ac · Pc (2) is established. 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 differential pressure between the pressure Pp and the pressure Pz is If it is constant, it can be seen that the movement amount x is determined by qs.

Further, if the opening area of the variable throttle 45 of the pilot spool valve 400 is a, the pilot flow rate qs
Passes through the opening area a, qs = C2 · a · (Pc−Pz) ^1/2 (6) where C2: the flow coefficient of the variable throttle 45 C2 · a · {K · Pp + (1-K) Pz-Pz} 1/2 = C2 · a · K 1/2 · (Pp-Pz} 1/2 ... (7) (7) equation (5) Substituting into the equation, 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 moving amount x of the seat valve element 20 is equal to the pilot spool valve 4 provided on the pilot line.
It is controlled by the opening area a of the variable stop 45 of 00.

On the other hand, the main flow rate flowing from the feeder passage 7C to the feeder passage 23 through the auxiliary variable throttle 28 of the sliding portion 20c of the seat valve 300 is represented by Qs,
Assuming that the outer diameter of c is L, the auxiliary variable aperture 2 of the sliding portion 20c
Since the opening area of 8 is the product of the outer diameter L and the moving amount x, Qs = C3 · L · x · (Pp−Pz) ^1/2 (9) where C3: the flow coefficient of the variable throttle 28 By substituting equation (5) 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)

Next, the pilot spool valve 4 shown in FIG.
At 00, the preset force of the spring 47 is applied to the spool 41 as an urging force in the valve opening direction, and the pilot pressure P2a or P2b, which is the operation signal of the main spool valve 200B of the second direction switching valve device 100B, is applied to the pressure receiving chamber 50. Is applied so as to act in the valve closing direction. For this reason,
Assuming that the pressure conversion value of the preset force of the spring 47 is F, the pressure conversion value of the spring constant of the spring 47 is K, the pilot pressure P2a or P2b is Pi, and the moving amount of the pilot spool 41 in the valve closing direction is X, the pilot spool 41 Is expressed as Pi = F + K.X (13). That is, the moving amount X of the pilot spool 41 is determined by the pilot pressure Pi, and when the pilot pressure Pi increases, the moving amount X of the pilot valve body 41 increases.
Increases, and the opening area of the pilot variable diaphragm 45 decreases.

Therefore, as described above, the seat valve body 2
Since the moving amount x of 0 is controlled by the opening area of the pilot variable throttle 45, the flow rate Qv of the pressure oil flowing from the feeder passage 7C to the feeder passage 7A or 7B can be controlled by the pilot pressure P2a or P2b. That is, the seat valve 300, the pilot lines 24, 29 to 31, 35
To a pair of main variable throttles 1 from the pump passage 5 through the feeder passage 7 by the pilot spool valve 400.
The flow rate of the pressure oil supplied to 6A, 16B is restricted according to the pilot pressure P2a or P2b (flow rate restriction signal), and the flow rate of the pressure oil flowing into the pair of load passages 6A, 6B is controlled in an auxiliary manner. .

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 moves in the valve closing direction. The auxiliary variable aperture 28 is fully closed, and the control variable aperture 33 is also fully closed. Therefore, the backflow of the pressure oil from the feeder passage 7A or 7B to the feeder passage 7C is prevented, and the seat valve 300 performs a load check function.

As described above, according to the present embodiment, in the first direction switching valve device 100A, the auxiliary flow rate is controlled by the combination of the seat valve 300, the pilot lines 24, 29 to 31, 35 to 37, and the pilot spool valve 400. The control function and the load check function are performed, and the following effects are obtained.

First, the first directional control valve device 100
Since A has an auxiliary flow rate control function, it is possible to auxiliaryally control the supply flow rate to only the intended flow control valve in a combined operation in which a plurality of actuators are simultaneously driven, and the combined operability is improved.

That is, in FIG. 2, the hydraulic control valve device 100 of this embodiment is used as a hydraulic drive 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 an arm. Arm cylinder 702 having a low load pressure and a swing motor 7 having a high load pressure at the time of starting.
Consider the case of performing the simultaneous operation No. 01. In this case, the pressure oil from the hydraulic pump 700 is supplied from the pump port 5a to the main spool valve for arm 200A and the main spool valve for turning 200B.
The pilot pressure P2a or P2 of the turning main spool valve 200B is supplied to the pilot spool valve 400 of the first direction switching valve device 100A at the same time.
b is given as a flow rate limiting signal, and the seat valve 300 moves the seat valve body 20 in the throttle direction in accordance with the pilot pressure P2a or P2b, and the main variable throttle 16A or 1A.
Control is performed so as to reduce the flow rate of the pressure oil supplied to 6B. For this reason, the pressure of the pressure oil supplied to the turning main spool valve 200B increases, and a necessary amount of pressure oil is supplied to the turning motor 701, so that an appropriate combined operation as intended by the operator can be performed.

The seat valve 300 is connected to the main spool valve 20.
It is installed in the position where the load check valve was located in the conventional valve device on the feeder passage 7 of 0A. Therefore, the seat valve 300 functions to supplementarily control the supply flow rate only to the target main spool valve 200A, and has no influence on the other main spool valves 200B and 200C. Therefore, when the actuator 701 and the actuator 703 are driven at the same time, a combined operation can be performed as usual. In addition, the installation of the seat valve 300 does not restrict the arrangement of the spool valve, and there is an effect that the degree of freedom in design is improved.

Second, the hydraulic control valve device 100 of this embodiment
In the first directional switching valve device 100A, only two valves, the seat valve 300 and the main spool valve 200A, are arranged in the feeder passage 7 and the load passages 6A and 6B constituting the main circuit. Actuator with reduced pressure loss and less energy loss when pressurized oil passes through the main circuit than a conventional hydraulic control valve device in which three valves, a valve, a load check valve, and a pressure compensating valve, are arranged in the main circuit. Operation becomes possible.

Third, in the conventional hydraulic control valve device provided with a pressure compensating valve, it was necessary to form a large 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 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, compared with the conventional hydraulic control valve device provided only with the load check valve without the pressure compensation function, the circumference of the balance piston and the balance piston itself are enlarged, and the valve block is in the axial direction of the balance piston, that is, at right angles to the main spool. Direction, and the outer shape of the valve block becomes larger. Also, the production of the valve block becomes complicated.

In the present embodiment, the seat valve 300 is disposed in place of the load check valve at the position of the feeder passage 7 where the load check valve was provided in the conventional hydraulic control valve apparatus without a pressure compensation function. The valve 400 can be arranged using the fixed block 2 that holds the seat valve body 20 separate from the housing 1. For this reason, the height L of the portion of the housing 1 where the seat valve 300 is located is the height of the portion where the load check valve of the conventional valve device without pressure compensation function is located (the second and third direction switching valve devices). 100B, 100C load check valve 770,
(The height of the portion where the 771 is located), and the overall size of the housing 1 can be reduced. Also, the fixed block 2 can be made smaller by arranging the pilot spool 41 in parallel with the main spool 4A. Therefore, the entire valve device can be made compact, which is advantageous in terms of cost, and can increase the degree of freedom in mounting on a construction machine to be used.

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 this embodiment, the hydraulic control valve device 100 provided with the spool valve 200A, which is a spool type flow control valve, can provide an auxiliary flow control function with high control accuracy. Further, the pressure loss does not increase in spite of the auxiliary flow control function, and the actuator can be driven with a small energy loss. In addition, the housing becomes compact, mounting on a construction machine becomes easy, and manufacturing becomes easy, so that the manufacturing cost of the valve device can be reduced. Furthermore, since the auxiliary flow rate control function is provided while using the center bypass type flow rate control valve, the supply flow rate of only the intended flow rate control valve can be supplementarily controlled in a combined operation in which a plurality of actuators are simultaneously driven. Operability can be improved.

Second Embodiment 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 7 are denoted by the same reference numerals. This embodiment further facilitates the production of a controllable variable throttle and makes the seat valve perform a load check function.

8 and 9, the first direction switching valve device 101A of the hydraulic control valve device 101 according to the present embodiment has a seat valve 301, and the seat valve body 20 of the seat valve 301 has a seat valve 301 as shown in FIG. A passage 121 is formed instead of the passage 29 shown in FIG.
In this passage 121, a check valve 122 for allowing the flow of the pressure oil from the feeder passage 7C toward the hydraulic chamber 24 and preventing the flow in the opposite direction is arranged. The pilot flow groove 31A formed in the seat valve body 20 has a land portion so that the control variable throttle 33A slightly opens when the seat valve body 20 is in the valve closing position, as shown by FC in FIG. 32 is set.

In the first embodiment shown in FIGS. 1 to 7, 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 control variable throttle 33 formed in the pilot flow groove 31 at this time is also fully closed, and the seat valve 300 performs a load check function. However, when the seat variable valve element 20 moves from the fully closed position in the valve opening direction, if the control variable throttle 33 does not immediately open, the pilot flow immediately after the opening becomes unstable. For this reason, in the configuration of the first embodiment, the positional relationship between the upper end of the pilot flow groove 31 and the land portion 32 is accurately determined so that the controllable throttle 33 is immediately opened when the seat valve element 20 moves in the valve opening direction. Must be processed.

On the other hand, in the present embodiment, when the seat valve body 20 moves to the fully closed position as described above, the upper end of the pilot flow groove 31A is closed so that the control variable throttle 33A is not completely closed. The positional relationship with the land 32 is set. As a result, a stable pilot flow can be generated, the flow control accuracy can be improved, and the control variable throttle 33A can be easily manufactured.

In this embodiment, the check valve 12 is provided in the passage 121 in the seat valve body 20 which forms a part of the pilot line.
Since the control variable throttle 33A is slightly opened when the seat valve body 20 is in the valve closing position, slight leakage of pressure oil through the pilot line is completely prevented, since
A load check function with high liquid tightness can be obtained. Since the check valve 122 is arranged in the pilot line, the check valve 122 is moved from the feeder passage 7C to the feeder passage 7A.
Alternatively, the pressure loss of the main flow rate flowing to 7B does not increase.

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 line, 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.

Third Embodiment A third embodiment of the present invention will be described with reference to FIGS.
In the drawings, members that are the same as the members shown in FIGS. 1, 2, 4, 6, 8, and 9 are given the same reference numerals. In this embodiment, the supply pressure is applied to the pilot spool valve to selectively use the auxiliary flow control function.

Referring to FIGS. 11 and 12, the first directional control valve device 102A of the hydraulic control valve device 102 of this embodiment has a pilot spool valve 401, and a pilot spool 820 of the pilot spool valve 401 has A pressure receiving chamber 821 extending in the axial direction and opening to the pressure receiving chamber 51 is additionally formed, and a slidable piston 822 having one end abutting on the screw 46 is inserted into the open end side of the pressure receiving chamber 821. The pilot spool 820 has a radial passage 823 connecting the pressure receiving chamber 821 to the outlet passage 43.
Is formed in the pressure receiving chamber 821.
The pressure in the feeder passage 7A or 7B is introduced through the passages 36, 37, 43, and 823, and the pressure is applied in the valve opening direction of the pilot spool 820.

In the present embodiment configured as described above, the seat valve 301 functions as follows in combination with the pilot spool valve 401.

As in the first embodiment, the pressure conversion value of the preset force of the spring 47 is F, the pressure conversion value of the spring constant of the spring 47 is K, the pivoting pilot pressure P2a or P2b is Pi, and the pilot spool 820 is Assuming that the amount of movement in the valve closing direction is X and the urging force by the pressure of the feeder passage 7A or 7B introduced into the pressure receiving chamber 821 is Fz, the balance of the force applied to the pilot spool 820 is according to the first embodiment. As in the above equation (13), Pi = F + K.X + Fz (14). That is, the moving amount X of the pilot spool 820 is determined by the pilot pressure Pi and the urging force Fz. If the pilot pressure Pi increases, the moving amount X of the pilot spool 41 increases and the pilot variable throttle 45
When the urging force Fz increases, the pilot spool 820 moves in the valve opening direction, and the movement amount X
Decreases, and the opening area of the pilot variable diaphragm 45 increases.

Therefore, in the combined operation example in which the swing motor 701 and the arm cylinder 702 are simultaneously driven as described above, when the load pressure of the arm cylinder 702 is high, the pilot spool 820 automatically moves in the valve opening direction to change the pilot variable throttle. 45, the moving amount x of the seat valve element 20 of the seat valve 301 is increased,
Unnecessary energy loss during actual excavation can be avoided.

Therefore, according to the present embodiment, the auxiliary flow rate control function can be exhibited only when the load pressure of the actuator 702 is low, and unnecessary energy loss can be avoided and economical efficiency can be improved while ensuring a good composite operation. .

Fourth Embodiment A fourth embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 1, 2, 4, 6, 8, and 9 are given the same reference numerals. In this embodiment, a control signal is introduced into the pilot spool valve as a flow rate limiting signal instead of the pilot pressure of another flow control valve.

In FIG. 13, reference numeral 500 denotes a pilot pump, and a relief valve 501 is connected to a discharge line 500a of the pilot pump to maintain a constant pressure in the pilot line 502. The pilot line 502 is connected to the primary side of an electromagnetic proportional pressure reducing valve 504, and the secondary side of the electromagnetic proportional pressure reducing valve 504 is connected to a pilot line 505.
Is connected to the passage 800 of the pilot spool valve 400. The electromagnetic proportional pressure reducing valve 504 includes a control device 506.
, And generates a control pressure Pc corresponding to the control signal. The control pressure Pc is used as a flow rate limiting signal via the line 505 and the passage 800 to control the pressure receiving chamber 5.
0 is introduced. The control device 506 receives a setting signal from a setting device 507 operated by an operator, and creates a control signal based on the setting signal.

FIG. 14 shows the configurations of the control device 506 and the setting device 507. The control device 506 has an input unit 506a, a calculation unit 506b, a data unit 506c, and an output unit 506d. The setting device 507 is a turning priority switch 507
a and an arm priority switch 507b.

In the above-described combined operation of simultaneously driving the turning motor 701 and the arm cylinder 702, if the user wants to give priority to turning, the operator gives the turning priority switch 507.
a is turned on, and a turning priority signal is set as a setting signal to the control device 5.
06 is output. The control device 506 inputs the turning priority signal via the input unit 506a, calculates the flow control amount using the turning priority signal and the data stored in the data unit 506c in the calculating unit 506b, and outputs a corresponding control signal. Part 5
06d to the electromagnetic proportional pressure reducing valve 504. The electromagnetic proportional pressure reducing valve 504 is controlled by the control signal from the control device 506 to generate a corresponding control pressure Pc, and this control pressure Pc is used as a flow rate limiting signal as the pilot spool valve 4.
00 pressure receiving chamber 50. The pilot spool valve 400 uses the control pressure Pc to control the pilot variable throttle 45 as in the case of the pilot pressure Pi of the first embodiment.
Control the pilot flow rate. In this case, a relatively large control pressure P corresponding to the turning priority signal
c is generated, and the opening area of the pilot variable stop 45 is relatively reduced. As a result, the seat valve is also throttled relatively strongly, and the pressure of the pressure oil supplied to the main spool valve 200B for rotation is relatively increased, so that a combined operation with a relatively high turning speed and a turning priority can be performed.

If the operator wants to give priority to the arm, the operator turns on the arm priority switch 507b, and an arm priority signal is output to the control device 506, and a control signal corresponding to the electromagnetic proportional pressure reducing valve 504 is output, as described above. Control pressure P corresponding to pressure receiving section 50 of spool valve 400
c is introduced. In this case, the control pressure Pc is relatively small, and the opening area of the pilot variable throttle 45 is slightly reduced. As a result, the seat valve is opened relatively widely, the pressure of the pressure oil supplied to the main spool valve for turning 200B is slightly increased, and the turning speed is relatively slow, and the arm raising is relatively fast. It becomes possible.

As described above, according to the present embodiment, the degree of priority of turning or arm can be adjusted by the operator's intention, and the composite operability is further improved.

FIG. 15 shows another configuration example of the system for generating the control pressure. In FIG. 15, reference numeral 510 denotes a pilot pressure P2 which is an operation signal of the main spool valve 200B for turning.
a pilot valve device that generates a or P2b, and includes pilot valves 510a and 510b that generate pilot pressures P2a and P2b, respectively, according to the operation amounts. A shuttle valve 511 is connected to pilot lines of the pilot valves 510a and 510b, and a pressure detector 512 is connected to an output line thereof. To the control device 506, in addition to the setting signal from the setting device 507, a pilot pressure signal detected by the pressure detector 512 is input. The control device 506 controls the flow rate using an arithmetic expression stored in advance in the data section 506 according to the setting signal from the setting device 507 and the detection signal of the pilot pressure P2a or P2b which is the operation signal of the turning main spool valve 200B. Calculate the quantity and output the corresponding control signal. Therefore,
In this control system, the degree of seat turning or the priority of the arm is adjusted according to the operator's intention and the magnitude of the operation signal of the turning main spool valve 200B, and the combined operability is further improved.

Fifth Embodiment A fifth embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 1, 2, 4, 6, 8, and 9 are given the same reference numerals. In this embodiment, the pilot spool valve has a variable relief function.

In FIGS. 16 and 17, the first directional control valve device 103A of the hydraulic control valve device 103 of this embodiment has a pilot spool valve 403, and a bore 840 of the pilot spool valve 403 has a pressure receiving chamber 50 And entrance passage 4
2, an annular passage 841 is additionally formed, and the fixed block 2 has a passage 842 opening to the annular passage 841.
Are formed. A pressure receiving chamber 8 extending in the axial direction and opening to the pressure receiving chamber 51 is provided inside the pilot spool 843.
In addition, a slidable piston 8 having one end abutting on the screw 46 is provided at the open end of the pressure receiving chamber 844.
45 is inserted. Also, the pilot spool 84
3, a radial passage 846 that connects the pressure receiving chamber 841 to the passage 841 is formed. On the other hand, the passage 842 is
7, it is connected to the pump port 5. Therefore, passages 842, 841, 846 are provided in pressure receiving chamber 844.
, The supply pressure of the pump port 5 is introduced, and the pressure is applied in the valve opening direction of the pilot spool 843.

In this embodiment configured as described above, the combination of the seat valve 301 and the pilot spool valve 403 has the following functions.

As in the first embodiment, the pressure conversion value of the preset force of the spring 47 is F, the pressure conversion value of the spring constant of the spring 47 is K, the turning pilot pressure P2a or P2b is Pi, and the pilot spool 820 is Assuming that the amount of movement in the valve closing direction is X and the urging force by the supply pressure of the pump port 5 introduced into the pressure receiving chamber 844 is Fp, the balance of the force applied to the pilot spool 843 is as described in the first embodiment. As in the equation (13), Pi = F + K.X + Fp (15). That is, the moving amount X of the pilot spool 843 is determined by the pilot pressure Pi and the urging force Fp. If the pilot pressure Pi increases, the moving amount X of the pilot spool 41 increases and the pilot variable throttle 45
When the urging force Fp (pump port pressure) increases, the pilot spool 843 moves in the valve opening direction and the movement amount X decreases.
5, the opening area increases.

Therefore, in the combined operation example in which the swing motor 701 and the arm cylinder 702 are simultaneously driven, the pressure of the pressure oil supplied to the swing main spool valve 200B is increased by the throttle action of the seat valve 301, and When the pump port pressure rises until the pilot pressure Pi and the pump port pressure become equal to the pressure receiving area ratio between the pressure receiving chamber 50 and the pressure receiving chamber 844, the pilot spool 843 starts to move in the valve opening direction to increase the opening area and increase the seat valve. 30
1 to reduce the throttle action. For this reason, the pressure of the pressure oil supplied to the turning main spool valve 200B becomes a value corresponding to the turning pilot pressure Pi, and the driving pressure of the turning motor 701 can be adjusted according to the pilot pressure Pi.

As described above, according to the present embodiment, the driving pressure of the turning motor 701 can be adjusted according to the turning pilot pressure Pi, whereby the combined operability is further improved.

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

Sixth Embodiment A sixth embodiment of the present invention will be described with reference to FIGS. In the present embodiment to the ninth embodiment, the pressure compensation control is performed by using the differential pressure across the main variable throttle as its flow restriction signal in a valve device having a center bypass type flow control valve. In the figure, the same reference numerals are given to the same members as those shown in FIGS. 1, 2, 4, and 6, and the description is omitted.

In FIGS. 18 and 19, the hydraulic control valve device of the present embodiment is generally indicated by reference numeral 105, and this hydraulic control valve device 105 is supplied to a hydraulic actuator 701 as shown in FIG. First direction switching valve device 105A for controlling the flow of pressure oil, second direction switching valve device 105B for controlling the flow of pressure oil supplied to hydraulic actuator 702, hydraulic actuator 70
3 has a third directional control valve device 105C for controlling the flow of the pressure oil supplied to the third directional control valve 3.

The hydraulic control valve device 105 includes a housing 1 common to the first to third directional switching valve devices, and a first directional switching valve device 1 integrally mounted on the housing 1.
05A, and a first directional switching valve device 105A is incorporated in the housing 1 to constitute a center bypass type flow control valve.
1A and a seat valve 300 incorporated in the housing 1
And a pilot spool valve 405 incorporated in the fixed block 2 to constitute a pilot flow control valve.

The main spool valve 201A of the first embodiment is the same as that of the first embodiment except that the operation method is a manual operation.
A and the seat valve 300 is exactly the same as the seat valve 300 of the first embodiment, including the associated pilot line.

The pilot spool valve 405 is different from the pilot spool valve 400 of the first embodiment except that the pressure difference between the pair of main variable throttles 16A and 16B is introduced as a flow rate limiting signal. Is the same. That is, as shown in an enlarged manner in FIG.
41 is a spool portion 94 located on the side of the bore bottom 40a.
1a and a spool portion 9 located on the open end side of the bore 40.
41b and a small-diameter portion 941c located near the land portion 44
And an inclined portion 941d connecting the small diameter portion 941c and the spool portion 941a. The inclined portion 941d cooperates with the land portion 44 to form a pilot variable throttle 45 located between the inlet passage 42 and the outlet passage 43. As shown in FIG. 7, the variable aperture 45 changes the opening area from a predetermined minimum opening to a predetermined maximum opening in accordance with the amount of movement of the pilot spool 941.

The preset force of the spring 47 disposed between the screw 46 and the pilot spool 941 is equal to the target value of the pressure difference between the front and rear of the main variable throttles 16A and 16B of the meter-in of the main spool valve 201A, as will be described later. The compensation differential pressure is set, and the spring 47 functions as target compensation differential pressure setting means.

The pilot spool 941 has an outlet passage 4
The passages 52 and 53 that connect the pressure chamber 3 to the pressure receiving chamber 50 are formed. In the pressure receiving chamber 50, the feeder passage 23, the passages 36, 3
The pressure in the feeder passages 7A, 7B is introduced through the outlet passage 7 and the outlet passage 43 and the passages 52, 53, and the pressure is applied in the valve closing direction of the pilot spool 941. Also,
The fixed block 2 is formed with a passage 54 that opens to the pressure receiving chamber 51 and passages 57 and 58 that communicate with the load passages 6A and 6B via passages 55 and 56 formed in the housing 1. The passage 54 and the passages 57 and 58 are formed. A shuttle valve 59 for taking out the pressure on the high pressure side of the passages 57 and 58 to the passage 54 is disposed between them. The pressure on the high pressure side of the load passages 6A and 6B is introduced into the pressure receiving chamber 51 through the passages 55 and 56, the passages 57 and 58, the shuttle valve 59 and the passage 54, and the pressure is increased in the valve opening direction of the pilot spool 941. Applied.
Due to the configuration of the pressure receiving chambers 50 and 51, the pilot spool valve 405 uses the differential pressure across the main variable throttles 16A and 16B as a flow rate limiting signal to control the pilot flow rate flowing through a pilot line composed of the passages 29 to 31, 35 to 37 and the like. Control.

Referring back to FIG. 18, both ends of the main spool 4A protrude from the end surface of the housing 1, respectively. The left end of the main spool 4A is connected to an operation lever (not shown) via a plug 75, and the right end of the main spool 4A is connected to a centering spring mechanism 77 via a plug 76. The centering spring mechanism 77 is covered by a cover 81 attached to the housing 1.

In FIG. 19, the configuration of the second and third directional control valve devices 105B and 105C is the same as that of the conventional center bypass type flow control valve, and the directional control valve devices 100B and 100B of the first embodiment are different from each other. 100C and the main spool valves 201B and 201C are the same except that they are manually operated.

In the hydraulic control valve device 105 of the present embodiment configured as described above, the first directional control valve device 105
A seat valve 300 is similar to that of the first embodiment,
It operates according to the principle described in JP-A-8-501781. That is, the pilot flow grooves 3 formed in the seat valve body 20
Aperture area for the first land portion 32 (the controllable aperture 33
Opening area) is the movement amount (stroke) of the seat valve body 20
And the amount of movement of the seat valve body 20 is determined according to the pilot flow rate passing through the pilot flow groove 31 (the controllable throttle 33). The pilot flow rate is determined by the opening area of the variable throttle 45 of the pilot spool valve 405. As a result, the feeder passage 2C is moved from the feeder passage 2C through the auxiliary variable throttle 28 of the seat valve body 20.
3, the main flow rate is proportional to the pilot flow rate, and the main flow rate is determined by the opening area of the variable throttle 45 of the pilot spool valve 405.

In the pilot spool valve 405, the opening area of the variable throttle 45 is controlled so that the pressure difference across the main variable throttle 16A or 16B is changed as a flow rate limiting signal.

As described above, the seat valve 300 is connected to the pilot lines 24, 29 to 31, 35 to 37 (see FIG. 4),
In combination with the pilot spool valve 405, the flow rate of the pressure oil supplied from the pump passage 5 to the main variable throttle 16A or 16B via the feeder passage 7 is controlled by the main variable throttle 16A.
Alternatively, an auxiliary flow rate control function is performed in which the flow rate of the hydraulic oil flowing into the pair of load passages 6A and 6B is controlled by limiting the pressure oil according to the differential pressure (flow rate restriction signal) of 16B. Hereinafter, this will be described in more detail.

First, with respect to the seat valve 300, the above-described equations (1) to (12) hold as described in the first embodiment.

In the pilot spool valve 405, the preset force of the spring 47 as target compensation differential pressure setting means is applied to the pilot spool 941 as a biasing force in the valve opening direction, and the feeder passage 7A or 7
The pressure in B is applied in the pressure receiving chamber 50 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 to act in the valve opening direction. Therefore, the load pressure is set to PL, the pressure conversion value of the preset force of the spring 47 is set to F, and the pressure receiving chambers 50, 5
Assuming that the pressure receiving area of the pilot spool 941 is equal to the pressure in the feeder passage 7A or 7B, the pressure in the feeder passage 7A or 7B is equal to the pressure Pz in the feeder passage 23 of the seat valve 300. PL + F = Pz (16)

By modifying the equation (16), Pz-PL = F (17) If the opening area of the main variable throttle 16A or 16B provided on the main spool 4A is A, the seat valve 300 is The relationship between the flow rate when the passed flow rate Qv passes through the main variable throttle 16A or 16B and the differential pressure before and after is represented by Qv = C4 · A · (Pz−PL) ^1/2 (18)

Using the equations (12) and (17), (1)
By transforming equation 8), qs = C4 · A / (1 + α) · F ^1/2 (19) is obtained. By transforming equation (18) using equation (17), the following equation is obtained: Qv = C4 · A · F ^1/2 (20)

The above equation (20) indicates that the main spool valve 201
The flow rate Qv (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 of A 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 set by the preset force F from the above equation (17).

Therefore, the seat valve 300 is connected to the pilot lines 24, 29 to 31, 35 to 37 (see FIG. 4),
In combination with the pilot spool valve 405, the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B is determined by the differential pressure across the main variable throttle 16A or 16B (flow rate limiting signal).
And a differential pressure Pz-PL across the main variable throttle 16A or 16B at this time.
Springs 47 regardless of load or supply pressure fluctuations.
The pressure compensation control is performed so that the preset force F of the set value coincides with the set target compensation differential pressure.

As described above, the hydraulic control valve device 10 of this embodiment
Since the first directional switching valve device 105A of No. 5 has a pressure compensating function, the flow control accuracy of the first directional switching valve device 105A is improved, and the operability when shifting from the single operation to the combined operation is improved. improves.

That is, in FIG. 19, the control valve device 105 of this embodiment is used for a hydraulic circuit device of a hydraulic shovel,
It is assumed that the hydraulic actuator 701 is a swing motor for rotating and driving the swivel table, and the hydraulic actuator 702 is a boom cylinder for raising and lowering the boom. Let's consider the case of transition to compound operation. In this case, the swing motor 701 in a single swing operation
Is relatively low, the operation amount of the direction switching valve device 105A (movement amount of the main spool 4A) is relatively small, and the swing motor 701 is rotating at a very low speed. Is higher than the load pressure of the swing motor 701, the feeder passages 7 and 774 of the direction switching valve devices 105A and 105B of both actuators.
Are connected in parallel, so that more hydraulic 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 1
If the first directional control valve device 105A has the same configuration as the second directional control valve device 105B without the pressure compensation function described above in 05A, an unexpected increase in the speed of the swing motor 701 is started by the combined operation. You. Such behavior occurs when the second direction switching valve device 105B is operated to raise the boom during crane operation while suspending and turning a fishing load, and the supply flow rate to the turning motor 701 increases rapidly, The turning speed changes suddenly and becomes a very dangerous state.

In this embodiment, since the above-mentioned pressure compensation function is provided to the first directional control valve device 105A, the flow rate Qv passing through the main variable throttle 16A or 16B is obtained.
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.

The seat valve 300 performs the load check function, the height L of the housing 1 does not increase, and the structure is simple, as in the first embodiment.

Therefore, according to the present embodiment, the hydraulic control valve device 105 provided with the spool valve 201A which is a spool type flow control valve can provide an auxiliary flow control function (pressure compensation function) with high control accuracy. it can. Further, the pressure loss does not increase in spite of the auxiliary flow control function, and the actuator can be driven with a small energy loss. In addition, the housing becomes compact, mounting on a construction machine becomes easy, and manufacturing becomes easy, so that the manufacturing cost of the valve device can be reduced. Further, since the pressure compensation function is provided while using the center bypass type flow control valve, the flow control accuracy is improved, and the operability when shifting from the single operation to the combined operation can be improved.

Seventh Embodiment A seventh embodiment of the present invention will be described with reference to FIGS. In the drawings, members equivalent to those shown in FIGS. 8 and 9 and FIGS. 18 to 20 are denoted by the same reference numerals. This embodiment is a modification of the sixth embodiment similar to that of the second embodiment.
6A, a passage 121 is formed in the seat valve body 20 of the seat valve 301 in place of the passage 29 shown in FIG. 18, and the passage 121 allows the flow of the pressure oil from the feeder passage 7C to the hydraulic chamber 24, and Check valve 12 that blocks flow
2 are arranged. Further, as shown by FC in FIG. 10, the pilot flow groove 31A formed in the seat valve body 20 has a land so that the control variable throttle 33A is slightly opened when the seat valve body 20 is in the valve closing position. The positional relationship with respect to the unit 32 is set. The check valve may be installed anywhere on the pilot line.

According to the present embodiment, similar to the second embodiment,
A stable pilot flow can be generated, the flow control accuracy is improved, and the manufacture of the controllable throttle 33A is facilitated. Further, a slight leak of pressure oil from the pilot line is completely prevented, and a highly liquid-tight load check function can be obtained.

Eighth Embodiment An eighth embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 18 to 20 and 21 are given the same reference numerals. In the present embodiment, the preset force of the spring in the pilot spool valve can be adjusted from outside.

23 and 24, the first directional control valve device 107A of the hydraulic control valve device 107 of this embodiment has a pilot spool valve 406, and the open end of the bore 40 of the pilot spool valve 406 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. Both ends of the pilot spool 94 are provided between the adjuster screw 130 and the pilot spool 941 as in the sixth embodiment.
1 and a spring 47 in contact with the screw 130 are arranged,
The preset force of the spring is applied as an urging force in the valve closing direction of the pilot spool 941.

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 sets the target value (target compensation differential pressure) of the differential pressure across the main variable throttles 16A and 16B of the main spool valve 201A, and sets 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,
By adjusting the pressure compensation characteristics of the seat valve 301, the flow characteristics of the first direction switching valve device 107A can be adjusted.

Therefore, according to the present embodiment, optimal pressure compensation characteristics and flow characteristics are set according to the type of actuator driven by the first direction switching valve device 107A, the type of load thereof, and the like. Operability can be improved.

Ninth Embodiment A ninth embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 18 to 20 and 21 are given the same reference numerals. In this embodiment, a hydraulic pressure generating means is provided instead of a spring as the target compensation differential pressure setting means of the pilot spool valve, and the target compensation differential pressure can be adjusted by making the pressure introduced thereto variable. .

Referring to FIGS. 25 and 26, the first directional control valve device 108A of the hydraulic control valve device 108 of this embodiment has a pilot spool valve 407, and the pilot spool valve 407 is configured as follows. I have.

The fixed block 2 has a bottom 140a at one end.
And a bore 14 having the other end opened to the outer surface of the fixed block.
In the bore 140, a spool valve element (hereinafter, referred to as a pilot spool) 141 of a pilot spool valve 407 is slidably disposed. The bore 140 is also formed in parallel with the bore 3 of the main spool valve 201A as in the previous embodiment (see FIG. 1), and the pilot spool 141 is correspondingly arranged in parallel with the main spool 4A (see FIG. 18). ing.

An annular pressure receiving chamber 150 is formed adjacent to the bottom 140a of the bore 140, and 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. A pressure receiving chamber 154 communicating with the passage 153 is formed between the screw 146 and the pilot spool 141.

The pilot spool 141 is provided at the bottom 1 of the bore.
A spool portion 141a located on the 40a side;
0, a spool portion 141b located on the open end side,
It has a small diameter part 141c located near the land part 144, and an inclined part 141d connecting the small diameter part 141c and the spool part 141a. The inclined portion 141d cooperates with the land portion 144 to change the pilot spool 141 between the inlet passage 142 and the outlet passage 143 from a predetermined minimum opening to a predetermined opening degree as shown in FIG. A pilot variable aperture 145 that changes the opening area up to the maximum opening is formed.

Inside the pilot spool 141, a pressure receiving chamber 155 extending in the axial direction and opening to the bore bottom 140a side.
And a pressure receiving chamber 15 extending in the axial direction and opening to the pressure receiving chamber 154 side.
A slidable piston 157 having one end contacting the bore bottom 140a is inserted into the opening end side of the pressure receiving chamber 155, and a slide contacting one end with the screw 148 at the opening end side of the pressure receiving chamber 156. A movable piston 158 is inserted. The pilot spool 141 has a radial passage 15 that connects the pressure receiving chamber 155 to the outlet passage 143.
9 and a radial passage 160 connecting the pressure receiving chamber 156 to the passage 151 are formed. In the pressure receiving chamber 155, the feeder 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
Of the pilot spool 14
1 is applied in the valve closing direction. Passage 5 is provided in pressure receiving chamber 156.
5, 56, passages 57, 58, shuttle valve 59 and passage 5
The pressure on the high pressure side of the load passages 6A, 6B is introduced through the pressure passage 4, and the pressure is applied in the valve opening direction of the pilot spool 141. The pressure receiving chambers 155 and 156 have the same inner diameter, and the pistons 157 and 158 also have the same outer diameter.
5, 156 and the pistons 157, 159 have the same pressure receiving area.

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 spool 141, and the pressure introduced into the pressure receiving chamber 150 is applied in the valve closing direction of the pilot spool 141.

In the pressure receiving chamber 154, a spring 163 having one end in contact with the pilot spool 141 and the other end in contact with the screw 146 is disposed.
Numeral 3 is provided for absorbing vibration, and the urging force of the spring 163 on the pilot spool 141 is negligibly small.

Therefore, the difference between the hydraulic pressure in the valve opening direction due to the constant pressure introduced into the pressure receiving chamber 154 and the hydraulic 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. It acts as a biasing force instead of the preset force of the spring 47 as the setting 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. 25, 5
Reference numeral 00a 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 506A, generates a variable pressure Pc according to the control signal, and the variable pressure Pc is introduced into the pressure receiving chamber 150.

In this embodiment constructed as described above, the seat valve 301 (FIG. 26) functions as follows in combination with the pilot spool valve 407.

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 urging force due to the difference from the variable pressure Pc introduced into the pilot spool is Fh, the balance of the force applied to the pilot spool 141 is: PL + Fh = Pz (Equation 16) in the same manner as in the aforementioned equation (16) according to the first embodiment. 21) is represented by

By transforming this equation (21), Pz-PL = Fh (22) Using the above equation (12) and equation (22), the flow rate when passing through the main variable throttle 16A or 16B By transforming the above equation (18) representing the relationship between the pressure and the differential pressure before and after, qs = C4 · A / (1 + α) · Fh ^1/2 (23) is obtained. By transforming equation (18), Qv = C4.A.Fh ^1/2 (24) is obtained. That is, similarly to the sixth embodiment, the flow rate Qv passing through the main variable throttle 16A or 16B of the main spool valve 201A is determined by 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 before and after the main variable throttle at this time.
Is a value indicated by the urging force Fh from the above equation (22).

Therefore, also in the present embodiment, the seat valve 301 functions as an auxiliary flow rate control means for restricting the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B.
At this time, the differential pressure P across the main variable throttle 16A or 16B
z-PL is pressure-compensated so that the biasing force Fh matches the target compensation differential pressure set regardless of the change in the load pressure or the supply pressure. 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 adjusted by the control device 506A and the electromagnetic proportional pressure reducing valve 504.
By using such a method, it is possible to carry out easily and with good controllability. Therefore, the target compensation differential pressure can be more finely adjusted, whereby the flow rate Qv passing through the main variable throttle 16A or 16B of the main spool valve 201A is more appropriately controlled, and the operability of the actuator is further improved. Can be.

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 spool 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.

Tenth Embodiment A tenth embodiment of the present invention will be described with reference to FIGS. In this embodiment to the fourteenth embodiment, in a valve device provided with a closed-center type flow control valve, pressure compensation control is performed by using a differential pressure across the main variable throttle as its flow restriction signal. In the figures, FIG. 1, FIG. 2, FIG.
The same reference numerals are given to members equivalent to those shown in FIGS.

In FIGS. 27 and 28, the hydraulic control valve device of the present embodiment is generally indicated by the reference numeral 110. As shown in FIG. It has a plurality of direction switching valve devices including the direction switching valve devices 110A and 110B. Further, the hydraulic control valve device 110 is a housing 1 common to a plurality of direction switching valve devices.
And a fixed block 2 integrally attached to the housing 1 and provided for each of the plurality of directional control valve devices.
The first directional control valve device 110A has a main spool valve 200 incorporated in the housing 1 to constitute a closed center type flow control valve, a seat valve 300 incorporated in the housing 1, and a fixed block. 2 and a pilot spool valve 405 which is incorporated in the pump 2 and constitutes a pilot flow control valve.

The main spool valve 200 is configured as follows. A bore 3 is formed through the housing 1, and a main spool 4 of a main spool valve 200 is slidably inserted into the bore 3. A pump port 5a connected to a hydraulic source (not shown) is provided in the housing 1 (see FIG. 28).
And a load passage 6 having load ports 6a and 6b connected to an actuator (not shown).
A, 6B, and the load passages 6A, 6B that branch off from the pump passage 5.
A feeder passage 7 (7A, 7B, 7C) that can communicate with B is formed.

In the bore 3, annular feeder passages 8A and 8B, which form part of the feeder passages 7A and 7B, and load passages 6A and
6B, annular load passages 9A and 9B, and annular discharge passage 10 communicating with tank port 85 (see FIG. 28).
A, 10B are formed, and the feeder passage 8A and the load passage 9 are formed.
A and land portions 11A and 12A are formed between the load passage 9A and the discharge passage 10A, respectively, and between the feeder passage 8B and the load passage 9B and between the load passage 9B and the discharge passage 10B. The lands 11B and 12B are respectively formed. A load detection passage 12 for detecting a load pressure is formed near the center of the bore 3, and a load detection port 13 for extracting the load pressure detected in the load detection passage 12 to the outside is formed in the housing 1. Is formed.

Notches 14A, 14B and notches 15A, 15B are formed on the main spool 4. Notch 14
A forms a meter-in main variable throttle 16A located between the feeder passage 8A and the load passage 9A in cooperation with the land portion 11A. The variable throttle 16A controls the amount of movement of the main spool 4 to the right in the drawing. Accordingly, the opening area is changed from the fully closed position to a predetermined maximum opening degree. The notch 14B forms a meter-in main variable throttle 16B located between the feeder passage 8B and the load passage 9B in cooperation with the land portion 11B. The variable throttle 16B moves the main spool 4 to the left in the drawing. The opening area is changed from the fully closed position to a predetermined maximum opening degree in accordance with. The notch 15B is located on the land 12
In cooperation with B, a meter-out main variable throttle 17B located between the load passage 9B and the discharge passage 10B is formed, and the variable throttle 17B is fully closed according to the amount of movement of the main spool 4 to the left in the drawing. The opening area is changed from the position to a predetermined maximum opening. The notch 15A forms a meter-out main variable throttle 17A located between the load passage 9A and the discharge passage 10A in cooperation with the land 12A.
Changes the opening area from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the main spool 4 to the right in the figure.

The seat valve 300 is the same as the seat valve 300 of the first embodiment, and the pilot spool valve 405 and the associated pilot line are the same as the pilot spool valve 405 and the associated pilot line of the sixth embodiment. .

The configuration of the second direction switching valve device 110B and the other direction switching valve devices is the same as the configuration of the first direction switching valve device 110A.

In the hydraulic control valve device 110 according to the present embodiment configured as described above, the seat valve 300 has been described in the first embodiment as in the sixth embodiment (1) to (1).
(16)-(2) explained in the equation (12) and the sixth embodiment.
Equation (0) holds. That is, the first direction switching valve device 1
In the 1A0 seat valve 300, 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 movement amount (stroke) of the seat valve body 20. The amount of movement of the seat valve element 20 is determined according to the pilot flow rate passing through the pilot flow groove 31 (controllable throttle 33).
The pilot flow rate is determined by the opening area of the variable throttle 45 of the pilot spool valve 405. As a result, the main flow flowing from the feeder passage 7C to the feeder passage 23 through the auxiliary variable throttle 28 of the seat valve body 20 is proportional to the pilot flow, and the main flow is the opening area of the variable throttle 45 of the pilot spool valve 405. Is determined.

In the pilot spool valve 405, the opening area of the variable throttle 45 is controlled so that the differential pressure across the main variable throttle 16A or 16B is changed as a flow rate limiting signal.

As described above, the seat valve 300 is connected to the pilot lines 24, 29 to 31, 35 to 37 (see FIG. 4),
In combination with the pilot spool valve 405, the flow rate of the pressure oil supplied from the pump passage 5 to the main variable throttle 16A or 16B via the feeder passage 7 is controlled by the main variable throttle 16A.
Alternatively, an auxiliary flow rate control function is performed in which the flow rate of the hydraulic oil flowing into the pair of load passages 6A and 6B is controlled by limiting the pressure oil according to the differential pressure (flow rate restriction signal) of 16B.

The seat valve 300 performs the load checking function, the height L of the housing 1 does not increase, and the structure is simple, as in the first and sixth embodiments.

Therefore, according to the present embodiment, in the hydraulic control valve device 110 provided with the spool valve 200 which is a spool type closed center type flow control valve, the auxiliary flow control function (pressure compensation function) with high control accuracy is provided.
Can be given. Also, despite the pressure compensation function and load check function, the pressure loss does not increase,
The actuator can be driven with a small energy loss. In addition, the housing becomes compact, mounting on a construction machine becomes easy, and manufacturing becomes easy, so that the manufacturing cost of the valve device can be reduced.

Eleventh Embodiment An eleventh embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 8 and 9 and FIGS. 27 and 28 are given the same reference numerals. This embodiment is a modification of the tenth embodiment in the same manner as the second embodiment.
First direction switching valve device 111A of hydraulic control valve device 111
, The check valve 122 is provided in the passage 121 of the seat valve 301.
Is installed. Further, as shown by FC in FIG. 10, the pilot flow groove 31A formed in the seat valve body 20 has a land so that the control variable throttle 33A is slightly opened when the seat valve body 20 is in the valve closing position. The positional relationship with respect to the unit 32 is set. The check valve may be installed anywhere on the pilot line. According to the present embodiment, similarly to the second embodiment, a stable pilot flow can be generated, the flow rate control accuracy is improved, and the manufacture of the controllable throttle 33A is facilitated. Further, a slight leak of pressure oil from the pilot line is completely prevented, and a highly liquid-tight load check function can be obtained.

Twelfth Embodiment A twelfth embodiment of the present invention will be described with reference to FIGS. In the drawings, members that are the same as the members shown in FIGS. 23 and 24, 27, 28, and 29 are given the same reference numerals. This embodiment is similar to the eighth embodiment except that the opening end of the bore 40 is closed by the adjuster screw 130 in the pilot spool valve 406 of the hydraulic control valve device 112. Is provided with an operation section 131 on the head thereof.

According to the present embodiment, similar to the eighth embodiment,
Optimum pressure compensation characteristics and flow characteristics can be set according to applications such as the type of actuator driven by the hydraulic control valve device 112 and the type of load thereof, and operability can be improved.

Thirteenth Embodiment A thirteenth embodiment of the present invention will be described with reference to FIG. In the drawing, members that are the same as the members shown in FIGS. 25 and 26, 27, 28, and 29 are given the same reference numerals. This embodiment is a modification of the ninth embodiment in the same manner as the ninth embodiment, except that a hydraulic pressure generating means is provided instead of a spring as the target compensation differential pressure setting means for the pilot spool valve. The target compensation differential pressure can be adjusted by making the introduced pressure variable.

That is, in FIG. 33, the directional control valve device 113A of the hydraulic control valve device 113 of this embodiment has a pilot spool valve 407, and this pilot spool valve 407 is the same as that of the ninth embodiment. It is configured similarly to the valve 407.

Further, in order to generate a constant pressure introduced into the pressure receiving chamber 154 and a variable pressure introduced into the pressure receiving chamber 150, the pilot pump 500 and the electromagnetic proportional pressure reducing valve 504 shown in FIG. , And a control device 506A.

According to this embodiment, the equations (21) to (24) hold as in the ninth embodiment, and the same effects as in the ninth embodiment can be obtained.

That is, also in this embodiment, the seat valve 3
01 functions as auxiliary flow rate control means for restricting the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B, and the differential pressure Pz− across the main variable throttle 16A or 16B at this time.
PL is controlled so as 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 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, which makes it possible to more appropriately adjust the main spool valve 2.
The flow rate Q passing through the main variable throttle 16A or 16B of 00
v can be controlled to further improve the operability of the actuator.

Fourteenth Embodiment A fourteenth embodiment of the present invention will be described with reference to FIGS. In the drawing, members that are the same as the members shown in FIGS. 27, 28, and 29 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 setting means for the pilot spool valve.

Referring to FIGS. 34 and 35, the hydraulic control valve device 114 of this embodiment has a first directional switching valve device 114A and a second directional switching valve device 114B. Main spool valve 204, seat valve 30
1 and the pilot spool valve 408.

That is, in FIG.
A main spool 221 of the main spool 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 5 having a pump port 5a, and load passages 6A and 6B branched from the pump passage 5 are provided.
And a feeder passage 7 (7A, 7B, 7C) that can communicate with the feeder.

In the bore 220, as in the embodiment shown in FIG. 27, annular feeder passages 8A and 8B, annular load passages 9A and 9B, and annular discharge passages 10A and 10B are formed. In between land parts 11A, 11B and 1
1B and 12B are formed respectively. Also, bore 2
The pump passage 5 is formed as an annular passage at the center of the pump 20, and the pump port 5 a of the pump passage 5 is connected to the hydraulic pump 600 (see FIG. 35).

The main spool 221 has notches 224A,
24B and notches 225A, 225B are formed. The notch 224A cooperates with the land 11A to form a meter-in main variable throttle 16A located between the feeder passage 8A and the load passage 9A.
Changes the opening area from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the main spool 221 to the right in the figure. The notch 224B cooperates with the land portion 11B to form a meter-in main variable throttle 16B located between the feeder passage 8B and the load passage 9B. The variable throttle 225B moves the main spool 221 to the left in the drawing. The opening area is changed from the fully closed position to a predetermined maximum opening degree in accordance with. Further, the notch 225B cooperates with the land portion 12B to form the load passage 9.
B and a discharge passage 10B, a meter-out main variable throttle 17B is formed. The variable throttle 17B moves from the fully closed position to a predetermined maximum opening in accordance with the rightward movement of the main spool 221 in the figure. Change the opening area. Notch 225
A forms a meter-out main variable throttle 17A located between the load passage 9A and the discharge passage 10A in cooperation with the land portion 12A.
The opening area is changed from the fully closed position to a predetermined maximum opening in accordance with the amount of movement of the left side of FIG.

Further, the feeder passage 7C and the feeder passage 7
A, 7B is connected to a valve body of the seat valve 301 (hereinafter, referred to as a valve body).
(Referred to as a seat valve element) 20 is disposed as appropriate. The configuration of the seat valve 301 is the same as that of the second embodiment shown in FIG. 29, and the description is omitted.

The land portions 11A and 11B have annular load detection chambers 230A and 230A for detecting 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 main spool 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.
The main spool 221 is provided at a position where the load pressure in the load passage 9B is taken out when the main spool 221 moves to the left in the drawing. Further, passages 232A, 233A,
234A is formed, and the load detection chamber 230A and the load detection passage 231A communicate with the discharge passage 10A via these passages 232A, 233A, 234A when the main spool 221 returns to the neutral position, and detect the detected load pressure. Reduce to tank pressure. Similar passages are provided in the main spool 221 for the load detection chamber 230B and the load detection passage 231B. By reducing the load pressure detected when the main spool 221 is neutral as described above, it is possible to prevent a wasteful increase in the discharge pressure of the hydraulic pump when the main spool 221 is neutral when used in a load sensing type hydraulic drive device. .

On the other hand, a pilot spool valve 408 is incorporated in the fixed block 2. The structure of the pilot spool valve 408 is similar to that of the embodiment shown in FIGS. 25 and 33, and the structure is enlarged and shown in FIG.
In the figure, the same reference numerals are given to members equivalent to the members shown in FIG.

Referring to FIG. 36, a bore 240 is formed in the fixed block 2, and a spool (hereinafter, referred to as a pilot spool) 141 of a pilot spool valve 408 is slidably disposed in the bore 240.

As in the embodiment shown in FIG. 25, an annular pressure receiving chamber 150, an annular inlet passage 142,
An annular outlet passage 143, an annular passage 151, an annular passage 153, and a screw hole 148 are formed.
A screw 146 is attached to 48, and the open end of the bore 140 is closed. A pressure receiving chamber 154 communicating with the passage 153 is formed between the screw 146 and the pilot spool 141, and a weak spring 163 for preventing vibration is disposed in the pressure receiving chamber 154. A pilot variable throttle 145 is formed between the land 144 formed between the inlet passage 142 and the outlet passage 143 and the inclined portion 141d of the pilot spool 141. Further, the pressure receiving chamber 15
Another annular passage 239 is formed between the zero and the entrance passage 142.

Inside the pilot spool 141, pistons 157 and 1 extending in the axial direction and slidable toward the open end are provided.
The pressure receiving chambers 240 and 241 into which the pressure-receiving chambers 58 are inserted are formed, and the pressure receiving chambers 240 and 241 are respectively provided with radial passages 242 and 2.
It communicates with the passages 239 and 151 via 43.

In the fixed block 2, a passage 251 for connecting the pressure receiving chamber 150 to the feeder passages 7A and 7B via a passage 250 formed in the housing 1, and a passage 252 for connecting the passage 153 to 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 spool 141, and the load receiving chambers 154 are provided in the pressure receiving chambers 154.
B, passages 231A and 231B, passage 252 and passage 15
3, the pressure in the load passage 6A or 6B is introduced,
The pressure is applied to the pilot spool 141 in the valve opening direction.

In the fixed block 2, passages 253 and 254 for connecting the passage 151 to the pump passage 5, a passage 255 for communicating with the load detection passages 231A and 231B,
Passages 256 and 257 communicating with similar load detection passages of a directional control valve (not shown), and passages 25 communicating with passages 239
8, 259, 260 are formed, and the passage 260 and the passage 2
A shuttle valve 261 for extracting pressure on the high pressure side of the passages 255 and 256 to the passage 260 is disposed between the shuttle valve 261 and the passages 55 and 256. These passages 253 and 254 are provided in the pressure receiving chamber 241.
The supply pressure of the pump port, that is, the discharge pressure of the hydraulic pump is introduced through the passages 151 and 243, and the pressure is applied in the valve opening direction of the pilot spool 141. Further, passages 255, 256, 25
7. The maximum load pressure of the plurality of actuators is introduced via the shuttle valve 261, the passages 258, 259, 260 and the passages 239, 242, and the pressure is applied in the valve closing direction of the pilot spool 141.

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.

Further, the second direction switching valve device 114B comprises:
As shown in FIG. 35, the structure is substantially the same as that of the first direction switching valve device 114A, and the same members are denoted by the same reference numerals and description thereof will be omitted.

The hydraulic control valve device 1 configured as described above
FIG. 35 shows a circuit configuration of a hydraulic drive device in which 14 is used. In FIG. 35, reference numeral 600 denotes a variable displacement hydraulic pump, the displacement of which is controlled by a load sensing type regulator 601. The discharge pipeline 602 of the hydraulic pump 600 is connected to the pump port 5a of the hydraulic control valve device 114. 603, 604
Is a hydraulic actuator, and the first direction switching valve device 1
The load ports 6a, 6b of 14A are connected to the first actuator 603 via actuator lines 605A, 605B, and the second actuator 604 is connected to the load ports 6a, 6b of the second directional control valve device 114B. It is connected via 606B. Further, the first and second directional switching valve devices 114A,
The tank port 85 of 114 </ b> B is connected to the tank 607 via the tank port 85. 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 further introduced into the pressure receiving chambers 241, 240, respectively. .

The discharge pressure of the hydraulic pump 600 is introduced into the regulator 601 via the pilot line 608, and the maximum load pressure is guided via 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, pilot spool valve 408
In FIG. 27, the urging force due to the pressure difference between the pump discharge pressure introduced into the pressure receiving chamber 240 and the maximum load pressure introduced into the pressure receiving chamber 241 serves as target compensation differential pressure setting means of the tenth embodiment shown in FIG. Acting instead of the preset force of the spring 47, the seat valve 301 can be provided with a pressure compensation function and a load check function as in the tenth embodiment.

That is, the pressure receiving areas of the pressure receiving chambers 150 and 154 are the same, the pressure receiving areas of the pressure receiving chambers 240 and 241 are the same, and the feeder passages 7A and 7B are similar to the tenth embodiment.
The internal pressure and the load pressure are Pz and PL, respectively, the discharge pressure of the hydraulic pump 600 is Pp, and the maximum load pressure is PLS.
Assuming max, the balance of the force applied to the pilot spool 141 is as described in the above (13) according to the tenth embodiment.
Similarly to the equation, Pp + PL = Pz + PLSmax (25)

By transforming this equation (25), Pz-PL = Pp-PLSmax = Fd (26) Therefore, the differential pressure between the pump discharge pressure and the maximum load pressure becomes the urging force Fd of the flow rate setting means.

Also, the above equation (12) and this (26)
(18) indicating the relationship between the flow rate when passing through the main variable throttle 16A or 16B and the differential pressure before and after using the equation.
By transforming the equation, the relationship between the pilot flow rate qs and the urging force Fd is expressed as qs = C4 · A / (1 + α) · Fd ^1/2 (27) By transforming equation (18) using equation (26), the flow rate Q supplied from the pump port to the load port
v is represented by Qv = C4 · A · Fd ^1/2 (28) That is, as in the first embodiment, the flow rate Qv of the main spool valve 204 passing through the main variable throttle 16A or 16B is determined by the urging force Fd 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 before and after the main variable throttle at this time.
Is a value corresponding to the urging force Fd from the above equation (26).

Therefore, also in the present embodiment, the seat valve 301 functions as an auxiliary flow rate control means for restricting the flow rate of the pressure oil supplied to the main variable throttle 16A or 16B, and before and after the main variable throttle 16A or 16B at this time. Differential pressure Pz
-PL is pressure-compensated so as to be equal to the target compensation differential pressure set by the urging force Fd regardless of fluctuations in the load pressure or the supply pressure. That is, the seat valve 301 can have a pressure compensation function and a load check function.

In the present 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 114A and 114B is controlled by the load sensing control of the hydraulic pump 600. When the discharge flow rate of the hydraulic pump 600 is insufficient during the combined driving of the hydraulic actuators 603 and 604, the pump discharge pressure and the maximum load are set. The differential pressure from the 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.

[0210]

According to the present invention, in a hydraulic control valve device and a hydraulic drive device used for a construction machine such as a hydraulic excavator, a spool type flow control valve which has a high reliability and is easy to design has been used for many years. In addition, an auxiliary flow rate control function with high control accuracy can be provided without increasing the pressure loss or increasing the size of the structure.

In the hydraulic control valve device and the hydraulic drive device provided with the center bypass type flow control valve,
In a combined operation in which a plurality of actuators are simultaneously driven, a supply flow rate to only a target flow control valve can be auxiliary controlled, and combined operability can be improved.

Furthermore, in a hydraulic control valve device and a hydraulic drive device provided with a center bypass type flow control valve,
It has a pressure compensation function and can improve composite operability.

Further, in the hydraulic control valve device and the hydraulic drive device provided with the closed center type flow control valve, it is possible to have a pressure compensation function and to avoid an increase in pressure loss and an increase in size of the structure.

[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 a diagram showing the opening degree characteristics of a seat valve and a controllable variable throttle shown in FIG. 4;

FIG. 6 is an enlarged view of a pilot spool valve in the hydraulic control valve device shown in FIG.

FIG. 7 is a view showing an opening degree characteristic of the pilot variable throttle shown in FIG. 6;

FIG. 8 is a sectional view of a hydraulic control valve device according to a second 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 diagram showing opening degree characteristics of the seat valve and the controllable throttle shown in FIG. 8;

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

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

FIG. 13 is a cross-sectional view of a pilot spool valve portion of a hydraulic control valve device according to a fourth embodiment of the present invention and a schematic diagram of a system for generating a flow rate limiting signal thereof.

14 is a diagram showing a configuration example of a system for generating a flow rate restriction signal of the embodiment shown in FIG.

FIG. 15 is a diagram showing another configuration example of the system for generating the flow rate restriction signal of the embodiment shown in FIG.

FIG. 16 is a sectional view of a hydraulic control valve device according to a fifth embodiment of the present invention.

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

FIG. 18 is a sectional view of a hydraulic control valve device according to a sixth embodiment of the present invention.

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

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

FIG. 21 is a sectional view of a hydraulic control valve device according to a seventh embodiment of the present invention.

FIG. 22 is a circuit diagram of a main part of the hydraulic control valve device shown in FIG.

FIG. 23 is a sectional view of a hydraulic control valve device according to an eighth embodiment of the present invention.

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

FIG. 25 is a cross-sectional view of a pilot spool valve portion of a hydraulic control valve device according to a ninth embodiment of the present invention and a circuit diagram of a related circuit configuration thereof.

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

FIG. 27 is a sectional view of a hydraulic control valve device according to a tenth embodiment of the present invention.

FIG. 28 is a circuit diagram of the hydraulic control valve device shown in FIG. 27.

FIG. 29 is a sectional view of a hydraulic control valve device according to an eleventh embodiment of the present invention.

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

FIG. 31 is a sectional view of a hydraulic control valve device according to a twelfth embodiment of the present invention.

FIG. 32 is a circuit diagram of the hydraulic control valve device shown in FIG. 31.

FIG. 33 is a circuit diagram of a hydraulic control valve device according to a thirteenth embodiment of the present invention.

FIG. 34 is a sectional view of a hydraulic control valve device according to a fourteenth embodiment of the present invention.

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

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

[Explanation of symbols]

100; 101-114 Hydraulic control valve device 100A; 101A; 102A; 103A; 105A;
106A; 107A; 108A; 110A; 111A;
112A; 113A; 114A Direction switching valve device 200A; 201A; 200; 204 Flow control valve 300; 301 Seat valve 400; 401; 403; 405; 406; 407;
08 Pilot control valve 1 Housing 2 Fixed block 3 Bore 4A; 4; 221 Main spool 5a Pump port 5 Pump passage 6a, 6b Load port 6A, 6B Load passage 7 (7A, 7B, 7C) Feeder passage 16A, 16B Main variable throttle Reference Signs List 20 seat valve element 21 bore 23 feeder passage (connection point) 24 hydraulic chamber (pilot line) 25 spring 29, 30 passage (pilot line) 28 auxiliary variable throttle 31 pilot flow groove (pilot line) 33 control variable throttle 35, 36, 37 Pilot passage (pilot line) 40 Bore 41; 820; 843; 941; 141 Pilot spool 45 Pilot variable throttle 47 Spring (first biasing means) 50, 51 Pressure receiving chamber (second biasing means) 122 Non-return Valve 130 Adjuster screw (operation Operating means) 131 Operating part (operating means) 150, 154 Pressure receiving chamber (first biasing means) 155, 156 Pressure receiving chamber (second biasing means) 240, 241 Pressure receiving chamber (first biasing means) 504 Electromagnetic proportional pressure reducing valve 506 Control device 507 Setting device Pc Control signal 500 Pilot pump (control signal generating means) 504 Electromagnetic proportional pressure reducing valve (control signal generating means) 505 Pilot line (control signal introducing means) 506 Control device (control signal generating means) ) 507 setting device (setting signal output means) 512 pressure detector 701-703 hydraulic actuator 800; 52-59; 159, 54-59; 231A,
231B, 251, 252 Passage (input means) 802 Shuttle valve (signal generation / transmission means) 803 Line (signal generation / transmission means) 821 Pressure receiving chamber (first biasing means) 844 Pressure receiving chamber (first biasing means) P1a , P1b Pilot pressure (flow rate limit signal)

──────────────────────────────────────────────────続 き Continuing from the front page (72) Inventor Masami Ochiai 650, Kandamachi, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura Works (56) References JP-A-50-61734 (JP, A) JP-A-4 -54303 (JP, A) JP-A-4-244605 (JP, A) JP-A-4-244604 (JP, A) JP-A-2-134402 (JP, A) (58) Fields investigated (Int. . ^6, DB name) F15B 11/00 E02F 9/22 F15B 11/05

Claims (20)

(57) [Claims] 1. A housing, a pump passage formed in the housing, and at least one direction switching valve means incorporated in the housing, wherein the direction switching valve means includes a pair of main variable throttles. A main spool slidably disposed in the housing to form a flow control valve, and a feeder passage formed in the housing to supply pressure oil from the pump passage to the pair of main variable throttles. And a pair of load passages formed in the housing and through which the pressure oil that has passed through the pair of main variable throttles respectively flows. An auxiliary flow rate that restricts a flow rate of the pressure oil supplied to the pair of main variable throttles through the feeder passage and controls a flow rate of the pressure oil flowing into the pair of load passages; Further comprising control means,
(A) a seat valve disposed in the feeder passage, the seat valve body being movably disposed in the housing, and forming an auxiliary variable throttle in the feeder passage; formed in the valve body, have a and aperture control variable to vary the opening area in accordance with the amount of movement of said seat valve body and said
The main spool is moved from the intermediate standing position down the main variable of said pair
Ri when in the state, but the open, low to prevent backflow of pressure oil
A seat valve which have a de check function; (b) the upstream side of the auxiliary variable throttle of said feeder passage through said control variable throttle to contact the downstream side of said feeder passage, the flow rate of the hydraulic fluid flowing through it pilot line and which determines the amount of movement of said seat valve body; comprising an input means for inputting a pyro Ttosupuru and the flow regulating signal (c) having a pilot variable throttle disposed in said pilot line, the flow rate limit signal Accordingly before
Serial wherein by moving the pilot spool pilot <br/> changing the opening area of the bets variable throttle, and the pilot spool valve <br/> for controlling the flow rate of the hydraulic fluid flowing through said pilot line; wherein the sheet The valve element is arranged orthogonal to the main spool.
Is, the pilot spool of the pilot spool valve
A hydraulic control valve device, which is arranged so as to be parallel to the main spool . 2. The hydraulic control valve device according to claim 1,
The direction switching valve unit further includes a fixed block that holds the seat valve body in the housing via a spring,
The hydraulic control valve device, wherein the pilot spool valve is incorporated in the fixed block. 3. The hydraulic control device according to claim 1, wherein the feeder passage is located upstream of the auxiliary variable throttle and communicates with the pump passage, and the auxiliary variable throttle of the feeder passage is provided. Downstream of the throttle, on both sides of the first passage portion, and second and third passage portions respectively communicating with the pair of main variable throttles, wherein the seat valve is provided in the first passage portion. A hydraulic control valve device, wherein the hydraulic control valve device is disposed at a connection point between the first and second passage portions. 4. The hydraulic control valve device according to claim 1,
The opening degree characteristic is set so that the control variable throttle is slightly opened at a fully closed position of the seat valve, and the direction switching valve means is further disposed on the pilot line and further includes a check valve for preventing a backflow of pressure oil. A hydraulic control valve device comprising: 5. The hydraulic control valve device according to claim 4 ,
The said non-return valve is built in the said seat valve body, The hydraulic control valve apparatus characterized by the above-mentioned. 6. The hydraulic control valve device according to claim 1 , further comprising a plurality of spool-type directional switching valve means incorporated in the housing, at least one of which is provided with the auxiliary. A hydraulic control valve device, which is a direction switching valve means having a flow control means. 7. The hydraulic control valve device according to claim 1, wherein
The hydraulic control valve device according to claim 1, wherein the input means of the pilot spool valve has a passage for inputting a hydraulic signal generated outside the direction switching valve means as the flow rate limiting signal. 8. The hydraulic control valve device according to claim 1, wherein
The pilot spool valve is connected to a first urging unit that applies a predetermined urging force to the pilot spool in a valve opening direction, and the input unit, and applies an urging force corresponding to the flow rate restriction signal to the pilot spool. A second biasing means for applying the force in the valve closing direction. 9. The hydraulic control valve device according to claim 8 , wherein
The hydraulic control valve device according to claim 1, wherein the first urging means includes a spring for urging the pilot spool in a valve opening direction with a predetermined preset force. 10. The hydraulic control valve device according to claim 9 , wherein said pilot spool valve further includes an operating means for externally adjusting a preset force of said spring. 11. A hydraulic control valve device according to claim 8 , wherein said first urging means has at least one pressure receiving chamber for applying a predetermined hydraulic pressure to said pilot spool in a valve opening direction. Hydraulic control valve device. 12. The hydraulic control valve device according to claim 8 , wherein the second urging means includes at least one pressure receiving chamber for applying a hydraulic pressure in the valve closing direction based on the flow rate restriction signal to the pilot spool. A hydraulic control valve device comprising: 13. The hydraulic control valve device according to claim 8 , wherein the input means introduces a hydraulic signal generated outside the direction switching valve means as the flow rate restriction signal to the second urging means. A hydraulic control valve device having a passage. 14. A hydraulic control valve device according to claim 13 , wherein said first urging means has a pressure receiving chamber into which inlet pressures of said pair of main variable throttles are introduced. Valve device. 15. The hydraulic control valve device according to claim 13 , wherein said first biasing means has a pressure receiving chamber into which the pressure of said pump passage is introduced. 16. A hydraulic pump; a plurality of hydraulic actuators driven by hydraulic oil discharged from the hydraulic pump; and a flow rate of the hydraulic oil supplied to the hydraulic actuators, each of which is operated in accordance with an operation signal. At least first and second directional switching valve means provided with a spool type flow control valve for controlling the
2. The hydraulic control valve device according to claim 1, wherein the direction switching valve means is a direction switching valve means having the auxiliary flow rate control means; and generating the flow rate limiting signal outside the first direction switching valve means. And a signal generating / transmitting means for introducing the signal into the input means of the pilot spool valve . 17. A hydraulic drive system according to claim 16 , wherein said signal generation and transmission means detects an operation signal given to said second direction switching valve means, and uses said operation signal as said flow rate restriction signal. Means for introducing to the input means of the pilot spool valve . 18. A hydraulic drive device according to claim 16 , wherein said signal generation and transmission means is operated by an operator to output a setting signal, and means for generating a control signal corresponding to said setting signal. Means for introducing a control signal into said input means of said pilot spool valve as said flow rate limiting signal. 19. A hydraulic drive device according to claim 16 , wherein said signal generation and transmission means is operated by an operator to output a setting signal, and an operation signal supplied to said second direction switching valve means and said setting signal. A hydraulic drive device comprising: means for generating a control signal corresponding to the signal; and means for introducing the control signal as the flow rate limiting signal to an input means of the pilot spool valve . 20. The hydraulic drive device according to claim 16 , wherein the flow control valve is a center bypass type spool valve.