JP2542005B2 - Road sensing control hydraulic drive - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a hydraulic drive system for a hydraulic machine including a plurality of hydraulic actuators such as a hydraulic excavator and a hydraulic crane. More specifically, the present invention relates to a hydraulic drive system using a load sensing regulator. Load sensing control hydraulic drive that controls the flow rate of the pressure oil supplied to the hydraulic actuator by the flow control valve device with pressure compensation while maintaining the discharge pressure of the pump at a pressure higher than the maximum load pressure of those hydraulic actuators by a predetermined value. Regarding the device.

[Conventional technology]

2. Description of the Related Art Conventionally, in a hydraulic machine including a plurality of hydraulic actuators such as a hydraulic excavator and a hydraulic crane, a load-sensing control hydraulic drive device is used in order to accurately control the hydraulic actuator and save energy. This hydraulic drive system generally includes a flow control valve device with pressure compensation connected between a hydraulic pump and a tank and each hydraulic actuator, and a discharge pressure of the hydraulic pump that is higher than a maximum load pressure of the plurality of hydraulic actuators. It is provided with a load sensing regulator that holds a predetermined value higher. The flow control valve device with pressure compensation guides the flow control valve that controls the flow rate of the pressure oil supplied to the hydraulic actuator and the inlet pressure and the outlet pressure of this flow control valve to the opposing pilot chambers, respectively, to load the hydraulic actuator. The pressure compensating valve is configured to keep the differential pressure before and after the flow control valve constant against fluctuations in pressure to compensate fluctuations in the flow rate.

The load sensing regulator, however, saves energy by setting the discharge pressure of the hydraulic pump to the minimum necessary value commensurate with the maximum load pressure of those hydraulic actuators, and at the same time, it is necessary for pressure compensation control of the flow control valve device with pressure compensation. A minimum differential pressure before and after that is ensured, and pressure compensation control of the flow rate of the pressure oil supplied to the hydraulic actuator is reliably performed.

However, in this load sensing control hydraulic drive device, a horsepower control device is further provided to the hydraulic pump, and when the engine driving the hydraulic pump is about to stall, the horsepower control device functions in preference to the load sensing regulator, Consumed flow rate of multiple hydraulic actuators, such as when reducing the discharge rate of the hydraulic pump or when the required flow rate of multiple hydraulic actuators commanded by the operating means exceeds the capacity (maximum discharge rate) of the hydraulic pumps. When the operating condition is such that the total of the discharge pressure of the hydraulic pump is greater than the discharge flow rate of the hydraulic pump, the load sensing regulator does not function and the discharge pressure of the hydraulic pump cannot be maintained higher than the maximum load pressure. Most of the pump discharge flow flows to the actuator,
There is a problem that pressure oil is not supplied to the high-pressure side hydraulic actuator, and the work element cannot be controlled in a desired manner.

Therefore, in order to solve such a problem, West German Patent Publication No. 3422165 discloses that the discharge pressure of the hydraulic pump is introduced to one of the pilot chambers facing the pressure compensating valve of the flow control valve device with pressure compensation, and the other To the maximum load pressure of multiple hydraulic actuators, and when the total flow rate of consumption of multiple hydraulic actuators is smaller than the discharge flow rate of the hydraulic pump, the load sensing regulator keeps the discharge pressure higher than the maximum load pressure by a specified value. As a result, if normal pressure compensation is performed and the total consumed flow rate becomes larger than the discharge flow rate, the throttle opening of each pressure compensation valve is reduced at the same rate as the discharge pressure decreases, and the flow control valve Distributing the discharge flow rate of the hydraulic pump according to the opening ratio so that the supply flow rate to each hydraulic actuator is reduced at the same rate. Is proposed. This is generally called non-saturation control of the discharge flow rate of a hydraulic pump.

On the other hand, in recent years, in hydraulic drive devices for hydraulic machines, it has been strongly desired to increase the hydraulic control pressure for the purpose of reducing the size and weight of hydraulic equipment.

[Problems to be solved by the invention]

The above-mentioned conventional example is satisfactory in that the discharge flow rate of the hydraulic pump is non-saturated in the normal control mode. However, the conventional hydraulic drive system is not suitable for increasing the hydraulic control pressure. That is, the flow control valve and the pressure compensating valve forming the flow control valve device with pressure compensation of this conventional example are both spool valves, and therefore, when high pressure acts on these valves, even if the valves are closed, A non-negligible liquid leak occurs from the gap that is inevitably provided to move the spool to and from the valve housing. For this reason, for example, when suspending a heavy load with a boom of a crane and holding it at that position and holding it at that position, if the hydraulic actuator is stopped with a large load applied, The hydraulic oil in the hydraulic actuator decreases due to liquid leakage from the hydraulic actuator, and the hydraulic actuator cannot be reliably stopped at that position.Therefore, situations such as hanging loads may occur, ensuring work safety. There was a problem that I could not do it.

In this prior art example, the flow control valve of the flow control valve device with pressure compensation is a directional control valve. Therefore, connecting a check valve between the flow control valve device with pressure compensation and the hydraulic actuator in order to prevent the liquid leakage is
This is impossible due to the circuit configuration.

Further, in this conventional example, since the pressure compensating valve is connected in series with the flow control valve in the main circuit, the pressure compensating valve must be of a size that allows the pressure oil in the main circuit to flow. However, the size of the pressure compensation valve becomes large, the entire hydraulic drive system becomes large, and the manufacturing cost thereof becomes high.

Therefore, an object of the present invention is to provide a load-sensing control hydraulic drive device that can perform non-saturation control of the discharge flow rate of a hydraulic pump, can handle high pressure of hydraulic control, and can be manufactured at low cost. That is.

[Means for solving problems]

The purpose of the above is to change the flow control valve device with pressure compensation by changing the opening degree according to the displacement of the inlet port and the outlet port connected to the main circuit between the hydraulic pump and each hydraulic actuator, and the valve body. A poppet valve having a back pressure chamber that communicates with the throttle and the inlet port via the variable throttle, and displaces the valve body in response to the differential pressure before and after the variable throttle, and the back pressure chamber and the outlet port of the poppet valve. Connected to the pilot circuit and the pilot circuit connected to the pilot circuit, the opening of the valve is changed by the operation of the operating means to change the pilot flow rate, and the differential pressure before and after the variable throttle in the poppet is changed to change the valve element. A pilot valve to be displaced and a pressure compensating valve having opposed pilot chambers connected to a pilot circuit between the back pressure chamber of the poppet valve and the pilot valve, and one of the pressure compensating valves Control pressure of the back pressure chamber of the poppet valve is independently guided to the pilot chamber, the maximum load pressure of multiple hydraulic actuators and the inlet pressure of the pilot valve are independently guided to the other pilot chamber, and the outlet of each poppet valve A check valve that blocks the flow of pressure oil toward the poppet valve and pilot circuit is connected to the main circuit downstream of the port and downstream of the pilot circuit, and the pressure receiving area at the inlet port of the poppet valve body is Ap, and the outlet is Let Aa be the pressure receiving area in the port, Ac be the pressure receiving area in the back pressure chamber, ac be the pressure receiving area of the one pilot chamber of the pressure compensating valve, and ar be the pressure receiving area of the other pilot chamber against the pilot valve inlet pressure. If the pressure receiving area for the maximum load pressure of the chamber is aam, the pressure receiving areas of the poppet valve body should be sized so that Ac = Aa + Ap. In the pilot chamber of the pressure compensating valve, the load sensing control hydraulic pressure is characterized in that the size of each pressure receiving area is determined so that when Ap / Ac = K is set, ar = (1-K) ac aamKac. Achieved by a drive.

[Action]

The control pressure in the back pressure chamber of the poppet valve is a value that changes in correlation with the discharge pressure of the hydraulic pump acting on its inlet port and the outlet pressure of the pilot valve at the same pressure as the load pressure acting on its outlet port. The pressure includes the effects of these discharge pressure and outlet pressure. Therefore, the pressure receiving area at the inlet port of the poppet valve body is Ap, the pressure receiving area at the outlet port is Aa, and the pressure receiving area at the back pressure chamber is
Let Ac be the ac pressure receiving area of the one pilot chamber of the pressure compensating valve, ac be the pressure receiving area of the other pilot chamber for the pilot valve inlet pressure, and aam be the pressure receiving area of the pilot chamber for the maximum load pressure. When the size of each pressure receiving area is determined so that Ac = Aa + Ap in the valve body and Ap / Ac = K is set in the pilot chamber of the pressure compensation valve, ar = (1-K) ac aam = By determining the size of each pressure receiving area so that it becomes Kac,
The control pressure can act as a substitute for the discharge pressure of the hydraulic pump or the outlet pressure of the pilot valve, and the control pressure of one pilot chamber of each pressure compensation valve and the pilot valve of the other pilot chamber The opposing action of the inlet pressure is functionally equivalent to the opposing action of the outlet pressure and the inlet pressure of the pilot valve. As a result, the differential pressure before and after the pilot valve is kept substantially constant with respect to the load fluctuation of the hydraulic actuator, the load pressure of the pilot flow rate is compensated, and the load pressure of the flow rate of the poppet valve is compensated.

Further, the fact that the control pressure guided to one pilot chamber and the maximum load pressure of the plurality of hydraulic actuators guided to the other pilot chamber act oppositely means that the discharge pressure of the hydraulic pump and the maximum load pressure oppose each other. It becomes functionally equivalent to that of the above, and as a result, when the total flow rate of consumption of multiple hydraulic actuators is about to exceed the discharge flow rate of the hydraulic pump, the pressure compensation valve The throttle opening is reduced at the same rate, the discharge flow rate of the hydraulic pump is distributed according to the rate of the corresponding opening of the pilot valve, and the flow rate supplied to the plurality of hydraulic actuators is reduced at the same rate. That is, non-saturation control of the discharge flow rate of the hydraulic pump is performed.

The check valve connected to the main circuit on the downstream side of the connection point between the outlet port of each poppet valve and the pilot circuit shuts off the flow of pressure oil from the hydraulic actuator to the poppet valve and the pilot circuit through the main circuit. For this reason, even if the hydraulic actuator is stopped under a heavy load, liquid leakage does not occur and it is safe. Therefore, it is possible to increase the pressure of the hydraulic drive device.

Further, since the pressure compensating valve is arranged not in the main circuit but in the pilot circuit with a small flow rate, the structure can be small.

〔Example〕

 Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows the overall circuit configuration of a load sensing control hydraulic drive system of the present invention. In the figure, reference numeral 1 is, for example, a swash plate type variable displacement hydraulic pump. A plurality of hydraulic actuators 2 and 3 are connected to the hydraulic pump 1, driven by pressure oil discharged from the hydraulic pump 1, and return oil from the hydraulic actuators 2 and 3 is directly returned to the tank. Between the hydraulic pump 1 and the hydraulic actuators 2 and 3, flow control valve devices 4 and 5 with pressure compensation for controlling the flow rate of the pressure oil supplied to the hydraulic actuators 2 and 3 are connected. The pump 1 is provided with a load sensing regulator 6 that holds the discharge pressure thereof higher than the maximum load pressure of the hydraulic actuators 2 and 3 by a predetermined value.

By using the load sensing regulator 6 in combination with the flow control valve devices 4 and 5 with pressure compensation, the discharge flow rate of the hydraulic pump 1 is suppressed to a necessary minimum to save energy, and the pressure of the flow control valve device with pressure compensation is reduced. The minimum differential pressure before and after that required for compensation control is secured, and hydraulic actuators 2, 3
Securely perform pressure compensation control of the flow rate of pressure oil supplied to
Stable control of the hydraulic actuators 2 and 3 can be performed.

Hereinafter, the detailed structures of the flow control valve devices 4 and 5 with pressure compensation and the load sensing regulator 6 will be described in this order. Since the pressure-compensated flow rate control valve devices 4 and 5 have substantially the same configuration, the constituent elements of the pressure-compensated flow rate control valve device 4 are denoted by a subscript a, and the pressure-compensated flow rate control valve device 5 has a subscript a. The constituent elements are given a subscript b, and only the former flow control valve device 4 will be described, and the description of the latter flow control valve device 5 will be omitted.

The pressure-compensated flow control valve device 4 includes a flow control valve 7a that controls the flow amount of the pressure oil supplied to the hydraulic actuator 2,
It comprises a pressure compensating valve 8a for compensating the fluctuation of the flow rate in the flow control valve 7a due to the fluctuation of the load pressure of the hydraulic actuator 2.

The flow rate control valve 7a is a poppet valve 11a as a main valve connected to the main pipelines 9a and 10a that form a main circuit, and a pilot flow rate is changed by changing an opening degree by operating an operating means. 11a and a pilot valve 12a for controlling the flow rate.

In the flow rate control valve 7a, the poppet valve 11a is configured as a flow rate amplification valve in which the valve opening is increased / decreased in proportion to the pilot flow rate flowing through the pilot valve 12a and a large flow rate obtained by amplifying the pilot flow rate is passed. Hereinafter, the configuration will be described in detail.

The poppet valve 11a has a valve housing 16a in which an inlet port 13a, an outlet port 14a and a pilot port 15a are formed, and the inlet port 13a is connected to the hydraulic pump 1 via a main pipe line 9a.
And the outlet port 14a is connected to the hydraulic actuator 2 via the main pipe line 10a. Pilot port 15
A pilot conduit 17a forming a pilot circuit is connected between a and the outlet port 14a, and the pilot valve 12a described above is connected to the pilot conduit 17a.

A valve seat 18a is formed in the valve housing 16a, and the valve body 19a
Are housed in such a manner that they can be moved in a sealed manner so as to come into contact with and separate from the valve seat 18a. The pilot port 15a communicates with the opposite side of the valve seat 18a in the valve housing. Disc 19a
A back pressure chamber 20a having a diameter larger than that of the back pressure chamber 20a is formed.
The back portion 21a of the valve body 19a located on the side farther from the valve seat 18a in 0a
Is arranged.

A plurality of grooves 22a are formed on the outer periphery of the valve body 19a.
The wall surface 38a that defines the upper end of a is a back pressure chamber at the valve closed position.
A variable throttle 23a that is positioned slightly above a wall surface 39a that defines the lower end of 20a, and that changes the opening degree according to the displacement of the valve body 19a is formed between the wall surfaces 38a and 39a. Therefore, the back pressure chamber 20a is always in communication with the inlet port 13a via the groove 22a and the variable throttle 23a.

With such a configuration, when the pilot valve 12a is closed and the pilot flow rate is not generated, the variable throttle 23
Since no pressure oil flows in a and no differential pressure is generated before and after the variable throttle 23a, the pressure in the back pressure chamber 20a becomes almost the same as the pressure in the inlet port 13a and acts on the lower end of the valve body 19a. Due to the difference between the pressure of the outlet port 14a and the pressure of the inlet port 13a acting on the upper end, a force for urging the valve body 19a in the valve closing direction acts, and the valve body 19a is closed.

When the operating means of the pilot valve 12a is operated and a pilot flow rate corresponding to the operation amount is generated in the pilot pipe line 17a, a flow of pressure oil is generated in the variable throttle 23a, which causes a differential pressure before and after the variable throttle 23a. Then, the pressure in the back pressure chamber 20a becomes smaller than the pressure in the inlet port 13a, and a force for urging the valve body 19a in the valve opening direction is generated, and the valve body 19a is moved in the valve opening direction. At this time, the opening degree of the variable throttle 23a increases as the displacement (opening degree) of the valve element 19a increases, and the differential pressure before and after that increases conversely, so the force that urges the valve element 19a in the valve opening direction. Is reduced in inverse proportion to the displacement of the valve body 19a, the force acting on the valve body 19a is balanced at a certain predetermined displacement,
The valve body 19a stops. That is, the back pressure chamber 20a displaces the valve body 19a in response to the differential pressure before and after the variable throttle.
In this way, the poppet valve 11a functions as a flow rate amplification valve that increases or decreases the valve opening in proportion to the pilot flow rate flowing through the pilot valve 12a and flows a large flow rate obtained by amplifying the pilot flow rate.

On the other hand, the pressure compensating valve 8a is connected to the pilot pipe line 17a between the pilot port 15a of the poppet valve 11a and the pilot valve 12a, and the pressure compensating valve 8a has first and second opposing members on both sides. A pilot having pilot chambers 31a and 32a, which displaces the spool between an open position and a closed position by pilot pressure acting on the first and second pilot chambers 31a and 32a and gives a throttle flow rate proportional to the spool displacement. It is configured as an operated two-position spool valve.

Since the pressure compensating valve 8a is thus arranged not in the main pipelines 9a and 10a but in the pilot pipeline 17a, it has a size enough to flow a pilot flow rate much smaller than the flow rate of the pressure oil in the main pipeline. It just needs to be made and it can be made extremely small.

The first pilot chamber 31a of the pressure compensation valve consists of two independent pilot chambers 36a and 37a, and one pilot chamber 3a
6a is connected to the pilot line 17a on the inlet side of the pilot valve 12a via the pilot line 33a, and the other pilot chamber 37a is connected to the maximum load pressure line 29 described later via the pilot line 35a. . The second pilot room 32
The a is connected to the pilot line 17a between the poppet valve 11a and the pressure compensation valve 8a via the pilot line 34a. Thereby, the first pilot chamber 31a of the pressure compensation valve 8a
Is independently guided to the maximum load pressure of the hydraulic actuators 2 and 3 and the inlet pressure of the pilot valve 12a, and the back pressure chamber 20a of the poppet valve 11a is introduced into the second pilot chamber 32a.
The control pressure at is independently derived.

Here, the pressure receiving area at the inlet port 13a of the valve body 19a of the poppet valve 11a is Ap (see FIG. 2, the same applies below), the pressure receiving area at the outlet port 14a is Aa, and the pressure receiving area at the back pressure chamber 20a is Ac, and pressure compensation is performed. The pressure receiving area of the pilot chamber 36a in the first pilot chamber 31 of the valve 8a is ar, and the pilot chamber 37
When the pressure receiving area of a is aam and the pressure receiving area of the second pilot chamber 32a is ac, the pressure receiving areas of the poppet valve body 19a are set to Ac = Aa + Ap (1). , In the pilot chambers 32a, 36a, 37a of the pressure compensating valve 8a, if Ap / Ac = K ... (2), then ar = (1-K) ac ... (3) aam = Kac ... (4) The size of each pressure receiving area is determined so that It should be noted that the equations (3) and (4) mean that the pressure receiving area has a relation of ac = ar + aam, as can be seen from a modification of these equations. That is, the ratio of the pressure receiving area ar, aam to the pressure receiving area ac of the pressure compensating valve 8a is the same as the ratio of the pressure receiving area Aa, Ap to the pressure receiving area Ac of the poppet valve body 19a, and the pressure receiving area ac, ar of the pressure compensating valve is , aam is the pressure receiving area of the poppet valve body
There is a relationship in which Ac, Aa, and Ap are copied in a similar manner.

As described above, the second pilot chamber 32a of the pressure compensation valve
The control pressure of the back pressure chamber 20a of the poppet valve 11a is introduced to the pilot pressure, which is the same as the discharge pressure of the hydraulic pump 1 acting on the input port 13a and the load pressure acting on the outlet port 14a. It is a value that changes in correlation with the outlet pressure of the valve 12a, and the control pressure includes the effects of these discharge pressure and outlet pressure. Therefore poppet valve body
By selecting the size of each pressure receiving area of 19a and the size of each pressure receiving area of the pilot chambers 32a, 36a, 37a of the pressure compensating valve 8a in a predetermined relationship, the control pressure is the discharge pressure of the hydraulic pump 1 or the pilot valve 12a. It is possible to function as a substitute for the outlet pressure of the pilot valve 32a of the pressure compensating valve and guide the inlet pressure of the pilot valve 12a to the pilot chamber 36a so that these pressures act in opposition. Is functionally equivalent to making the outlet pressure and the inlet pressure of the pilot valve face each other. As a result, the differential pressure before and after the pilot valve is kept substantially constant with respect to the load fluctuation of the hydraulic actuator, the load pressure of the pilot flow rate is compensated, and the load output of the flow rate of the poppet valve is compensated.

Further, the control pressure is guided to the pilot chamber 32a of the pressure compensating valve 8a, the maximum load pressure is guided to the pilot chamber 37a, and these pressures act in opposition so that the discharge pressure of the hydraulic pump 1 and the maximum load pressure act in opposition. It is functionally equivalent to doing. As a result, when the total flow rate of the hydraulic actuators 2 and 3 is about to exceed the discharge flow rate of the hydraulic pump 1, the throttle openings of the pressure compensating valves 8a and 8b are reduced according to the decrease in discharge pressure. It decreases at the same rate, and the hydraulic pump 1 distributes the discharge flow rate according to the ratio of the opening of the corresponding pilot valves 12a, 12b, and the plurality of hydraulic actuators
The supply flow rate to 2,3 is reduced at the same rate. That is, non-saturation control of the hydraulic pump 1 is performed.

The relationship between the pressure receiving areas of the above formulas (1) to (4) is determined by the size of each pressure receiving area of the poppet valve body 19a and the size of each pressure receiving area of the pilot chambers 32a, 36a, 37a of the pressure compensating valve 8a. The above-mentioned pressure compensation and non-saturation control functions are realized. This will be described in more detail later in the description of the operation of this embodiment.

In the main circuit downstream of the outlet port 14a of the poppet valve 11a and downstream of the pilot conduit 17a, that is, downstream of the pilot conduit connection point 24a of the main conduit 10a, from the hydraulic actuator 2 to the poppet valve 11a and the pilot conduit. A check valve 25a that blocks the flow of pressure oil toward the path 17a is connected.

A pilot line 26a is connected between the pilot line 17a on the downstream side of the pilot valve 12a and the maximum load pressure line 29, and the pressure in the pilot line 17a is connected to the maximum load pressure line 29a. Check valve that opens when it is higher than the pressure in line 29 and selects that pressure as the maximum load pressure
27a is connected. The maximum load pressure line 29 is
Is connected to the tank via a line 30 equipped with.

 Next, the load sensing regulator 6 will be described.

The load sensing regulator 6 is an auxiliary pump 40
A cylinder type swash plate tilting device 41 driven by the pressure oil discharged from the auxiliary pump 40, and an auxiliary pump
The control valve 42 is connected between the swash plate tilting device 41 and the swash plate tilting device 41. The control valve 42 is an auxiliary pump for the swash plate tilting device 41.
Between the first position communicating with the 40 and the second position communicating with the tank, the opening is continuously reduced according to the displacement of the spool, and the flow rate supplied to the swash plate tilting device 41 is controlled. Has become. In addition, the control valve 42 is provided on both sides of the
And second pilot chambers 43 and 44, and a spring 45 for setting pressure is provided on the side of the second pilot chamber 44.

The first and second pilot chambers 43 and 44 are connected to the main pipe line 9a on the discharge side and the maximum load pressure pipe line 30 via pilot lines 46 and 47, respectively. As a result, the discharge pressure of the hydraulic pump 1 is introduced into the first pilot chamber 43, and the two hydraulic actuators 2, 3 are introduced into the second pilot chamber 44.
The maximum load pressure of is derived.

When the maximum load pressure of the hydraulic actuators 2 and 3 rises, the control valve 42 causes a part of the pressure oil supplied to the swash plate tilting device 41 to escape to the tank, increasing the swash plate tilting amount of the hydraulic pump 1. Increase the discharge flow rate, resulting in an increase in discharge pressure,
Hold the amount of swash plate tilt when it balances with the maximum load pressure. When the maximum load pressure of the hydraulic actuators 2 and 3 decreases, the opposite operation is performed. As a result, the discharge pressure of the hydraulic pump 1 is maintained at a pressure higher than the maximum load pressure of the hydraulic actuators 2 and 3 by a predetermined value corresponding to the strength of the spring 45. In other words, the differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the hydraulic actuators 2 and 3 is maintained at a constant value set by the spring 45.

Next, the operation of the load sensing control hydraulic drive device configured as described above will be described.

First, when neither of the operating means of the pilot valves 12a, 12b is operated, the pilot valves 12a, 12b are closed and the pilot flow rate does not flow in the pilot conduit 17a.
Therefore, as described above, the variable throttle 2 of the poppet valves 11a and 11b is
No pressure difference is generated across 3a and 23b, the valve elements 19a and 19b are held in the closed position, and the poppet valves 11a and 11b are closed.

Therefore, pressure oil is not supplied to the hydraulic actuators 2 and 3,
These hydraulic actuators 2 and 3 are stopped.

Next, the case where the operating means of the pilot valves 12a and 12b are operated to individually drive the hydraulic actuators 2 and 3 will be described. For convenience of explanation, each state quantity in this hydraulic drive system will be expressed as follows.

Discharge pressure of hydraulic pump 1: Pp Higher load pressure of hydraulic actuators 2, 3 ie maximum load pressure; Pam Spring 45 set pressure: Ps Pilot valve 12a inlet pressure: P11 Pilot valve 12a outlet pressure: P12 Poppet valve Control pressure of 11a: Pc First, when the operation means of the pilot valve 12a is operated, the pilot valve 12a opens at an opening degree corresponding to the operation amount. As a result, the pilot flow rate according to the opening of the pilot valve 12a flows through the pilot pipe line 17a. As a result, as described above, the poppet valve is proportional to the pilot flow rate.
The valve opening of 11a is opened, and a large flow rate obtained by amplifying the pilot flow rate flows into the head side of the hydraulic actuator 2. Therefore, the hydraulic actuator 2 is driven at a speed according to the operation amount of the operation means of the pilot valve 12a.

At this time, the first pilot chamber 31 of the pressure compensation valve 8a
The pressure P11, Pam acts on the pilot chambers 36a, 37a of a via the pilot conduits 33a, 35a, and the pressure Pc, acts on the second pilot chamber 32a via the pilot conduit 34a. Since the spool of the pressure compensating valve 8a is balanced by these pressures, the relation of ac Pc = ar P11 + aam Pam (5) is established in consideration of the pressure receiving areas ac, ar, aam described above.

On the other hand, at this time, also in the poppet valve 11a, since the valve body 19a is pressure balanced, the pressure receiving area A
Considering p, Aa and Ac, the relation of Ac Pc = Ap Pp + Aa P12 (6) is established. Substituting Aa = Ac-Ap from Ac = Aa + Ap in the above-mentioned equation (1) into this equation (6), Ac Pc = Ap Pp + (Ac-Ap) P12 ... (7) is obtained. Since Ap / Ac = K is put in the equation (2), the equation (7) can be transformed by using this K. Is obtained. Further, in the expressions (3) and (4) described above, the relationship of ar = (1-K) ac and aam = Kac is defined. Therefore, in Pc, ar, aam of the above equation (5), equation (8),
Substituting the values of equations (3) and (4), Can be obtained. In this equation (9), K / (1-K) is a constant, so if this is put as C, equation (9) becomes P11-P12 = C (Pp-Pam) (10). Especially when K = 0.5, that is, poppet valve body 1
When the pressure receiving area Ap at the inlet port 13a of 9a is half of the pressure receiving area Ac of the back pressure chamber 20a (thus, the pressure receiving area Aa at the outlet port 14a is also half of Ac), C = 1 and P11-P12 = Pp-Pam ... (11)

That is, assuming that the pressure difference between the pilot valve 12a and the pilot valve 12a is ΔP, the pressure compensating valve 8a calculates the pressure difference ΔP by C (Pp−
Pam) to control.

On the other hand, at this time, in the hydraulic pump 1, the discharge pressure of the hydraulic pump 1 is the maximum load pressure Pam of the hydraulic actuators 2 and 3 due to the action of the load sensing regulator 6. In this case, since the hydraulic actuator 3 is not driven, the hydraulic pressure is reduced. The pressure is kept higher than the load pressure Pam of the actuator 2 by a predetermined value (set pressure) corresponding to the strength of the spring 45. As a result, Pp = Pam + Ps holds, and the following relation is obtained from this.

Pp-Pam = Ps = constant (12) Therefore, by comparing equations (10) and (12), P11-P12 = constant, and the differential pressure ΔP before and after the pilot valve 12a is kept constant. I understand. Therefore, the pilot flow rate flowing through the pilot valve 12a is also kept constant, and the large flow rate flowing through the poppet valve 11a is also kept constant. That is, pressure compensation is performed.

As a result, even if the load acting on the hydraulic actuator 2 changes and the differential pressure across the poppet valve 11a changes, the flow rate through the poppet valve is kept constant and the hydraulic actuator 2 has a pilot valve 12a. A desired flow rate according to the operation amount of the operating means is accurately supplied.

Then, when the operating means of the pilot valve 12a is returned to the neutral position, the pilot valve 12a is closed and the poppet valve 11a is also closed. This also stops the expansion of the hydraulic actuator 2.

When the hydraulic actuator 2 is extended in this way and stopped while the hydraulic actuator 2 is still under load, the high pressure oil on the head side of the hydraulic actuator 2 is fed by the check valve 25a to the poppet valve 11a and the pilot line. Backflow through 17a is prevented. Therefore, the pressure oil on the head side of the hydraulic actuator 2 is reduced, and the hydraulic actuator 2 is prevented from contracting due to the load. The poppet valve 11a itself is a valve that causes less liquid leakage than the spool valve originally. Therefore, even if the check valve 25a is not provided, as long as the check valve 25a is closed, the pressure oil does not leak from here and the pressure oil on the head side of the hydraulic actuator 2 does not decrease. However, the poppet valve 11a is closed when pressure oil is being discharged from the hydraulic pump 1. When the hydraulic pump 1 is stopped and pressure oil is no longer discharged, the valve body 19a of the poppet valve 11a is closed. The force that pushes in the direction disappears, and the poppet valve 11a is ready to open. By providing the check valve 25a, the pressure oil on the head side of the hydraulic actuator 2 will not flow out through the poppet valve 11a even if the poppet valve 11a is opened for this purpose. As a result, even if the hydraulic pump 1 is stopped due to an erroneous operation or the like, the pressure oil on the head side of the hydraulic actuator 2 is reliably prevented from decreasing, and safety is improved.

Therefore, for example, when the boom of the crane is operated by the hydraulic actuator 2 to lift and lift a heavy object and hold it at that position, the suspended load can be reliably stopped at that position, and the safety is improved. Can be secured.

When operating the operating means of the pilot valve 12b to operate the hydraulic actuator 3 alone, the operation is exactly the same as described above for the hydraulic actuator 2.

Next, a case where the operating means of the pilot valves 12a and 12b are simultaneously operated to simultaneously drive the two hydraulic actuators 2 and 3 will be described. For convenience of explanation, the state quantity in this hydraulic drive system is further defined as follows.

Inlet pressure of pilot valve 12b: P21 Outlet pressure of pilot valve 12b: P22 Discharge flow rate of hydraulic pump 1: Qp Consumed flow rate of hydraulic actuator 2: Q2 Consumed flow rate of hydraulic actuator 3: Q3 Operate means of hydraulic actuators 2 and 3 at the same time When operated, the poppet valve 1 for each of the flow control valves with pressure compensation 4 and 5, as in the case where these operating means are operated independently.
The opening of 1a and 11b opens in proportion to the pilot flow rate. As a result, on the head side of the hydraulic actuators 2 and 3, a large flow of pressure oil, which is the amplified pilot flow, is applied to the poppet valve 11 respectively.
Supplied through a, 11b. As a result, the hydraulic actuators 2 and 3 are driven at a speed according to the operation amount of the operation lever.

On the other hand, at this time, as described above, the pressures P11, Pam and Pc act on the pilot chambers 36a, 37a and the second pilot chamber 32a of the first pilot chamber 31a of the pressure compensating valve 8a, respectively. From the pressure balance, the relationship of P11-P12 = C (Pp-Pam) shown in the above equation (10) is obtained.

Also in the pressure compensating valve 8b of the flow control valve device 5 with pressure compensation, similarly, the pressure P21, Pb, is supplied to the pilot chambers 36b, 37b of the first pilot chamber 31b via the pilot conduits 33b, 35b.
Pam acts, and the pressure P'C acts on the second pilot chamber 32b through the pilot conduit 34b. From the pressure balance of the spool, the process similar to that of the above equation (10) is followed, and P21- The relationship of P22 = C (Pp-Pam) is obtained. Therefore, P11-P12 = C (Pp-Pam) and P21-P22 = C (Pp-Pa
From both equations of m), the relation of P11-P12 = P21-P22 = C (Pp-Pam) (13) is obtained.

That is, the pressure compensating valves 8a and 8b are the pilot valves 12a and 12b.
Let ΔP be the differential pressure before and after b, and the differential pressure ΔP
To C (Pp-Pam).

On the other hand, in such a state, when the sum of the flow rates consumed by the two hydraulic actuators 2 and 3 is smaller than the discharge flow rate of the hydraulic pump 1, that is, in the normal operating state of Q1 + Q2 ≦ Qp, the load sensing regulator 6 operates. Due to
The discharge pressure of the hydraulic pump 1 is maintained at a pressure higher than the maximum load pressure Pam of the hydraulic actuators 2 and 3 by a predetermined value (set pressure) corresponding to the strength of the spring 45, as described above (12).
The relations of expressions are obtained.

Therefore, by comparing equations (12) and (13),
P11-P12 = P21-P22 = constant, pilot valves 12a, 12
The differential pressure ΔP before and after b is kept constant. As in the case of operating the operating means of one pilot valve, the large flow rate of the poppet valves 11a and 11b is kept constant and pressure compensation is performed.

In such a state, two hydraulic actuators
The total of the consumption flow rates of 2 and 3 becomes larger than the discharge flow rate of the hydraulic pump 1. That is, assume that a situation occurs where Q1 + Q2> Qp. When such a situation occurs, as described above, the hydraulic pump 1 is further provided with a horsepower control device, and when the engine driving the hydraulic pump 1 is about to stall, the load sensing regulator 6 is prioritized. The horsepower control device functions to reduce the discharge flow rate of the hydraulic pump 1, and in addition to the two hydraulic actuators 2 and 3, there is a hydraulic actuator driven by the hydraulic oil discharged from the hydraulic pump 1. In the drive device, the total of the required flow rates instructed by the operating means may exceed the capacity of the hydraulic pump 1 (maximum discharge flow rate).

In such a case, the load sensing regulator 6 does not function, and as a transient state, the discharge pressure of the hydraulic pump 1 becomes smaller than the maximum load pressure, that is, Pp <Pam. Therefore, in equation (13) above, Pp−P
am <0 and P11-P12 = P21-P22 <0. Therefore, the pressure compensating valves 8a and 8b are both moved in the valve closing direction, and the flow rate of the pressure oil flowing through the pilot pipe lines 17a and 17b is adjusted to the pilot valve 12a
2b upstream to reduce the pilot flow rate through the pilot valves 12a, 12b. Therefore, the poppet valve 11a,
The flow rate flowing through 11b also decreases, and the flow rate supplied to the hydraulic actuators 2 and 3, that is, the total flow rate consumed by these actuators also decreases. Therefore, again, the total consumption flow rate of the hydraulic actuators 2 and 3 becomes smaller than the discharge flow rate of the hydraulic pump 1, and the pump discharge pressure becomes larger than the maximum load pressure. That is, Q1 + Q2 ≦ Qp and Pp> Pam.

Therefore, in the above equation (13), Pp-Pam> 0 again, and P11-P12 = P21-P22> 0. Therefore, the pressure compensating valve 8
The movement of a and 8b in the valve closing direction is suppressed, and the pilot pipelines 17a and
The reduction of the pilot flow rate flowing through 17b is prevented. Therefore, the flow rate of the hydraulic actuators 2 and 3 flowing through the poppet valves 11a and 11b is prevented from decreasing.

And finally, the discharge pressure of the hydraulic pump 1 is 0 <Pp.
The pressure reaches a predetermined value in the range of −Pam ≦ Ps, the pressures acting on the pilot chambers of the pressure compensation valves 8a and 8b are balanced in this state, and the pressure compensation valve stops at a predetermined throttle position. That is, Pp-Pam in P11-P12 = P21-P22 = C (Pp-Pam) of the above equation (13) becomes 0 <Pp-Pam≤Ps, and the pressure compensating valves 8a, 8
Both b are balanced when they are throttled to the same opening.

Therefore, the pilot flow rate flowing in the pilot pipes 17a, 17b also decreases while maintaining the ratio according to the ratio of the opening degree of the pilot valves 12a, 12b determined by the operation amount of each operation means, and flows through the poppet valves 11a, 11b. The flow rate supplied to the hydraulic actuators 2, 3 also decreases. As a result, the discharge flow rate of the hydraulic pump 1 is distributed according to the ratio of the opening degrees of the pilot valves 12a, 12b in the flow control valve devices with pressure compensation 4, 5, and the supply flow rate to each hydraulic actuator 2, 3 is the same. Decrease in proportion. That is, non-saturation control of the discharge flow rate of the hydraulic pump 1 is performed.

Finally, one of the hydraulic actuators 2 and 3, for example, the hydraulic actuator 2 is being driven, and a load is being applied to the other hydraulic actuator 3, while operating the pilot valve 12b with respect to the other hydraulic actuator 3. The operation when is operated will be explained.

First, when the hydraulic actuator 2 is driven and the load pressure of the hydraulic actuator 2 is P1a, the discharge pressure Pp of the hydraulic pump 1 is P1a + Ps due to the action of the load sensing regulator 6. This discharge pressure acts on the inlet side of the pilot valve 12b in the pilot line 17b via the main line 9b and the back pressure chamber 20b.

On the other hand, a load pressure is generated on the head side of the hydraulic actuator 3, and when the load pressure is P2a, this load pressure acts on the outlet side of the check valve 25b in the main pipeline 10b.

If the pilot valve 12b is opened to a predetermined opening by operating the operating means of the pilot valve 12b in such a state, when P2a> P1a, the discharge pressure of the hydraulic pump 1 is instantaneously affected by the action of the load sensing regulator 6. To Pp ＝
The pressure rises to P2a + Ps, and the hydraulic actuator 3 can be driven.

That is, immediately after opening the pilot valve 12b, there is no flow of pressure oil in the pilot line 17b, so the initial pump discharge pressure of Pp = P1a + Ps is transmitted as it is to the downstream side of the pilot valve 12b of the pilot line 17b. To be This pressure Pp
Is transmitted to the maximum load pressure line 29 via the check valve 27b, and Pp is selected as the maximum load pressure instead of P1a. As a result, the discharge pressure of the hydraulic pump 1 is increased to Pp + Ps, that is, P1a + 2Ps by the action of the load sensing regulator 6. This increased pressure is transmitted to the downstream side of the pilot valve 12b of the pilot line 17b, and the maximum load pressure line 29
This time, the increased pressure of P1a + 2Ps is transmitted, and the discharge pressure of the hydraulic pump 1 is further increased by the action of the load sensing regulator 6. This is repeated, and when the pressure on the downstream side of the pilot valve 12b in the pilot line 17b rises until it reaches the load output P2a of the second hydraulic actuator 3, the check valve 25b becomes openable, so the pilot valve 12b downstream. Side pressure does not rise above the load pressure P2a. That is, the pump discharge pressure rises by P2a + P
End with s. As a result, the pilot flow rate according to the opening of the pilot valve 12b flows through the pilot pipe line 17b, and the hydraulic actuator 3 can be driven.

As described above, according to the present embodiment, the flow rate control valve of each pressure compensation flow rate control valve device is configured by the poppet valve controlled by the pilot valve connected to the pilot line, and the flow rate control valve is connected to each pilot line. Incorporate a pressure compensating valve, and for each of the pressure compensating valves, independently guide the control pressure of the poppet valve to one of the opposing pilot chambers, and independently guide the inlet pressure and maximum load pressure of the pilot valve to the other. Therefore, when the sum of the flow rates consumed by the multiple hydraulic actuators is about to exceed the discharge flow rate of the hydraulic pump, the throttle opening of each pressure compensation valve is reduced at the same rate as the discharge pressure decreases, Distributing the discharge flow rate of the hydraulic pump according to the ratio of the opening of the corresponding pilot valve, reducing the supply flow rate to multiple hydraulic actuators at the same rate, In addition to the pressure compensation of the flow rate supplied to the hydraulic actuator by Le, it is possible to perform non-saturation control of the discharge rate of the hydraulic pump.

A check valve that blocks the flow of pressure oil toward the poppet valve and pilot line is connected to the main circuit on the downstream side of the outlet port of each poppet valve and on the downstream side of the pilot line. Even if the load is stopped in a state where the load is applied, the load can be surely stopped at a desired position without leakage of pressure oil, which is safe. Therefore, it is possible to increase the pressure of the hydraulic circuit device.

Further, since the pressure compensating valve is arranged not in the main circuit but in the pilot pipe line with a small flow rate, the structure can be made small, and the entire structure of the pressure compensating flow rate control valve device can be made small. Therefore, the flow control valve device with pressure compensation can be manufactured at low cost.

Further, the functions of the pressure compensation and the non-saturation control can be obtained by not introducing the control pressure of the poppet valve to the other pilot chamber of the pressure compensating valve alone, but also introducing the discharge pressure of the hydraulic pump and the outlet pressure of the pilot valve. Although the advantages of high pressure and miniaturization can be obtained, the number of pilot lines for the auxiliary valve is large, and therefore the piping becomes slightly complicated. Therefore, by guiding the control pressure of the poppet valve alone to the other pilot chamber of the pressure compensating valve as in the present embodiment, it is possible to simplify the piping and further integrate the poppet valve and the pressure compensating valve.

FIG. 2 is a specific example of a pressure compensating valve showing simplification of piping, and an enlarged view of a poppet valve is also shown therein. Further, in the figure, reference numerals a and b are added to indicate the pressure compensating valves 8a and 8b and the poppet valves 11a and 11b of the flow control valve devices 4 and 5 in common.
Is omitted.

The pressure compensation valve 8 has a valve housing 53 having an inlet port 50 and an outlet port 51 connected to the pilot line 17, and a pilot port 52 connected to the pilot line 35. A spool 54 for controlling the valve opening is housed. The spool 54 includes a first spool portion 55 having a pressure receiving area aam and a second spool portion 55 having a pressure receiving area ac.
And a rod portion 57 connecting the spool portions 55, 56.

Inside the valve housing 53, a first hydraulic pressure chamber 58 facing the outer end surface of the first spool portion 55, a second hydraulic pressure chamber 59 facing the outer end surface of the second spool portion 56, and a first hydraulic pressure chamber 59. Spool part
A third hydraulic chamber 60 facing the inner end surface of 55, a fourth hydraulic chamber 61 facing the inner end surface of the second spool portion 56, a second hydraulic chamber 59, and a third hydraulic chamber 60. And a passage 62 for communicating with the first hydraulic chamber 58, the first hydraulic chamber 58 communicates with the pilot port 52, the second hydraulic chamber 59 communicates with the inlet port 50, and the fourth hydraulic chamber 61.
Communicates with the outlet port 51. An annular valve seat portion 63 is provided around the rod portion 57 to define a flow passage having an area corresponding to the stroke of the spool 54 between the inner end surface of the first spool portion 55 and the valve seat portion 63. However, the flow rate corresponding to that is obtained. Further, the inner diameter of the valve seat portion 63 is determined so that an effective pressure receiving area ar that corresponds to an area portion between the outer diameter thereof and the inner diameter of the valve seat portion 63 is defined on the inner end surface of the second spool portion 56. Has been.

In the above-described configuration, the pilot flow rate flowing through the pilot pipe line 17 is introduced from the inlet port 50 into the second hydraulic pressure chamber 59, passes through the passage 62 and the third hydraulic pressure chamber 60, and flows through the valve seat portion 63 to the spool 54. While controlling the flow rate according to the stroke amount,
It flows out to the pilot line 17 again via the fourth hydraulic chamber 61 and the outlet port 51.

On the other hand, the maximum load pressure is introduced from the pilot line 35 into the first hydraulic chamber 58, and this maximum load pressure acts on the outer end surface of the first spool portion 55 having the pressure receiving area aam, and the spool 54 is , The pressure equal to the inlet pressure of the pilot valve 12 is introduced into the fourth hydraulic chamber 61, and this pressure has an effective pressure receiving area a r of the second spool portion 56. It acts on the inner end face to urge the spool 54 to the left. Further, a pressure equal to the control pressure of the poppet valve 11 is introduced into the second hydraulic chamber 59, and this pressure acts on the outer end surface of the second spool portion 56 having the pressure receiving area ac and the spool 5
Looking at 4 in the figure, it is biased to the right. That is, the inlet pressure and the maximum load pressure of the pilot valve 12 and the control pressure of the poppet valve 11 act on the spool 54 in opposition. As a result, pressure compensation and non-saturation control are performed as described above.

In this structure, the fourth hydraulic chamber 61 has the above-mentioned first
Of the pilot chamber 31a, the first hydraulic chamber 58 corresponds to the pilot chamber 37a, and the second hydraulic chamber 59 corresponds to the second pilot chamber 32a. That is, the second and fourth hydraulic pressure chambers 59 and 61 serve both as the flow passage for the pressure oil flowing through the pilot conduit 17 and the pilot chamber. Further, the pilot lines 33a and 34a shown in FIG. 1 are configured as internal passages,
The second and fourth hydraulic chambers 59 and 61 also have the role.

As described above, according to the configuration of this specific example, only one pilot conduit is required, and the piping can be simplified. Further, since the hydraulic chamber also serves as the flow passage for the pressure oil and the pilot chamber, the structure of the pressure compensating valve itself is simplified. Further, according to this pressure compensation valve, the poppet valve 11 is provided in the second hydraulic chamber 59.
It is also possible to omit the pilot line 17 between the pressure compensating valve 8 and the poppet valve 11 and to integrate the both, because in this case, the entire control pressure is introduced. The valve structure of is simple.

When the hydraulic drive system shown in FIG. 1 is constructed using the pressure compensating valve 8 shown in FIG. 2, the circuit configuration is as shown in FIG.

The embodiments of the present invention have been described above with reference to FIGS. 1 to 3, but the present invention is not limited to these embodiments, and various modifications and changes can be made within the scope of the spirit of the present invention. .
For example, in the above embodiments, the hydraulic pump 1 has been described as a variable displacement type, but it is not limited to this and may be a fixed displacement type hydraulic pump. In this case, the load sensing regulator can be configured as shown in FIG. That is, the load sensing regulator 80 has a pilot line in the pilot chamber on one side of the relief valve 81.
The discharge pressure of the fixed displacement hydraulic pump 82 is led via 46, the maximum load pressure is led to the pilot chamber on the other side where the spring 83 is arranged via the pilot line 47, and the set pressure of the relief valve 81 is set. Is determined by the spring 83 and the maximum load pressure, whereby the discharge pressure of the hydraulic pump 1 is kept higher than the maximum load pressure of the plurality of hydraulic actuators by a predetermined value corresponding to the spring 83.

Further, in the above embodiments, the poppet valves 11a and 11b forming the main valves of the flow control valves 7a and 7b are the valve bodies 19 having the illustrated shapes.
Although it has been described as having a and 19b, it is not limited to this, and it is connected to the variable throttle that changes the opening according to the displacement of the valve body and the inlet port through the variable throttle, Other configurations can be used as long as the configuration has a back pressure chamber that displaces the valve body in response to a differential pressure.

Further, although the above embodiments have been described with respect to the case where the two hydraulic actuators 2 and 3 are driven by the hydraulic pump 1, the present invention is not limited to this, and the present invention is also applied to a hydraulic drive device having three or more hydraulic actuators. It is possible.

The configuration of the pressure compensating load compensating valve devices 4 and 5 is not limited to the configuration shown in FIG. 2, and each valve as an appropriate element may be integrated, or the interlocking means of the pilot valve 12a and the check valve 24a may be mechanical. However, various modifications such as a hydraulic operation method or an electric operation method are possible.

〔The invention's effect〕

As is apparent from the above, in the load sensing control hydraulic drive device of the present invention, the flow control valve device with pressure compensation is
It consists of a flow control valve consisting of a poppet valve and a pilot valve, and a pressure compensating valve, and the poppet valve is placed in the main circuit.Therefore, non-saturation control of the discharge flow rate of the hydraulic pump can be performed and the main circuit downstream of the poppet valve can be controlled. Since the check valve is connected, liquid leakage of the hydraulic actuator does not occur, and even if the load pressure of the hydraulic actuator is high, the load can be safely held and the hydraulic control can be increased in pressure. Further, since the pressure compensating valve is arranged not in the main circuit but in the pilot circuit, the pressure compensating valve is downsized, and the hydraulic drive device can be manufactured at low cost.

[Brief description of drawings]

FIG. 1 is a circuit diagram schematically showing a load sensing control hydraulic drive system according to an embodiment of the present invention, and FIG.
FIG. 3 is a cross-sectional view showing a specific configuration of the flow control valve device with pressure compensation shown in FIG. 1, and FIG. 3 is a first diagram of a load sensing control hydraulic drive device using the flow control valve device with pressure compensation. 4 is a circuit diagram similar to FIG. 4, and FIG. 4 is a circuit diagram showing a configuration of a load sensing regulator when the load sensing control hydraulic drive device is realized by a fixed displacement hydraulic pump. In the figure, reference numeral 1 ... hydraulic pump, 2,3 ... hydraulic actuator, 4,5 ... pressure compensation flow control valve device, 6 ... load sensing regulator, 7a, 7b ... flow control valve, 8a,
8b ... Pressure compensation valve, 11a, 11b ... Poppet valve, 12a, 12b ...
… Pilot valves, 13a, 13b …… Inlet ports, 14a, 14b ……
Exit port, 17a, 17b …… Pilot line, 19a, 19b ……
Valve body, 20a, 20b ... Back pressure chamber, 23a, 23b ... Variable throttle, 24a,
24b …… throttle, 25a, 25b …… check valve, 31a, 31b …… first pilot chamber, 32a, 32b …… second pilot chamber, 3
3a, 33b, 34a, 34b, 35a, 35b …… Pilot line,

Claims (1)

(57) [Claims] 1. A flow control valve device with pressure compensation, which is connected between at least one hydraulic pump, a plurality of hydraulic actuators driven by pressure oil discharged from the hydraulic pump, and the hydraulic pump and each hydraulic actuator. And a load sensing regulator that keeps the discharge pressure of the hydraulic pump higher than the maximum load pressure of multiple hydraulic actuators by a predetermined value. In response to the differential pressure from the maximum load pressure of the actuators, when the total flow rate of consumption of multiple hydraulic actuators is about to exceed the discharge flow rate of hydraulic pumps, the supply flow rate to multiple hydraulic actuators should be kept at the same ratio. In a load-sensing control hydraulic drive configured to decrease the flow control valve device with each pressure compensation Through an inlet port and an outlet port connected to the main circuit between the hydraulic pump and each hydraulic actuator, a variable throttle for changing the opening degree according to the displacement of the valve body, and the variable throttle for the inlet port. A poppet valve having a back pressure chamber that communicates with and displaces the valve body in response to a differential pressure across the variable throttle, a pilot circuit connected between the back pressure chamber of the poppet valve and the outlet port, and this pilot. A pilot valve that is connected to the circuit and changes the opening by operating the operating means to change the pilot flow rate and changes the differential pressure before and after the variable throttle of the poppet valve to displace the valve element and the back pressure chamber of the poppet valve. And a pilot compensator connected to the pilot circuit between the pilot valve and a pressure compensating valve having opposite pilot chambers, and one of the pilot compensating valves of each pressure compensating valve has a control pressure of the back pressure chamber of the poppet valve. And independently guide the maximum load pressure of multiple hydraulic actuators and the inlet pressure of the pilot valve to the other pilot chamber, respectively, and to the downstream side of the outlet port of each poppet valve and the downstream side of the pilot circuit. Connect a poppet valve and a chuck valve that blocks the flow of pressure oil to the pilot circuit to the circuit.The pressure receiving area at the inlet port of the poppet valve body is Ap, the pressure receiving area at the outlet port is Aa, and the pressure receiving area in the back pressure chamber. Where Ac is the pressure receiving area of the one pilot chamber of the pressure compensating valve, ac is the pressure receiving area of the other pilot chamber for the pilot valve inlet pressure, and aam is the pressure receiving area of the pilot chamber for the maximum load pressure. In the poppet valve body, determine the size of each pressure receiving area so that Ac = Aa + Ap, and in the pilot chamber of the pressure compensating valve, If you place a Ap / Ac = K, ar = (1-K) ac aam = load sensing control hydraulic drive apparatus characterized by defining the size of each pressure receiving area so that Kac.