JP2007032782A - Hydraulic driving device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a hydraulic driving device compact in size and having improved operability at reduced cost achieved by eliminating a shuttle valve and a differential-pressure pressure reducing valve of the hydraulic driving device having a load-sensing system. <P>SOLUTION: The hydraulic driving device comprises a variable-displacement pump part, directional selector valves for a plurality of actuators, and pressure-compensation valves for compensating pressure of the directional selector valves. In the pressure compensation valves 22, 23, a spool 62 is slidably fit in a valve main body 60. The spool 62 has a structure where in the valve-closing direction, the upstream pressure Pin1 of the directional selector valve 18 acts on a first pressure-receiving area A21, and in the opening direction of the pressure-compensation valves, a load pressure PL1 of the actuator acts on a third pressurized area A23. Further, pressure Pr' acts on a second pressure-receiving area A22. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a hydraulic drive device used for construction machines such as a hydraulic excavator and various work machines. More specifically, the discharge pressure of a variable displacement hydraulic pump (hereinafter referred to as a variable pump) is a maximum load pressure of a plurality of actuators. The present invention relates to a hydraulic drive device that uses a load sensing system (hereinafter referred to as an LS system) that performs control so as to be higher by a target differential pressure.

  Conventionally, in this type of technology, an actual differential pressure (hereinafter referred to as Pc pressure) between the discharge pressure of the variable pump 2 and the maximum load pressure of the plurality of actuators 10 and 20 is detected by the differential pressure reducing valve 31 in the control valve 22. The hydraulic drive device is described. This hydraulic drive device has a plurality of actuators 10 and 20, and all of the compensation differential pressures of the pressure compensation valves 4 and 14 for controlling the differential pressure across the direction switching valves 8 and 18 for the actuators 10 and 20 are Pc pressures. It is controlled by. In this LS system, even when the plurality of actuators 10 and 20 are moved simultaneously, even when the required flow rate of the control valve 22 is in a saturation state exceeding the discharge flow rate of the variable pump 2, It has an anti-saturation function in which the flow rate is distributed at a ratio of 18 opening areas.

  This hydraulic drive device detects the Pc pressure by the differential pressure reducing valve 31 and feeds it back to a regulator valve (not shown) of the variable pump 2. The displacement of the variable pump 2 is made variable by balancing the Pc pressure and a spring (not shown) corresponding to the target compensation differential pressure and controlling the swash plate angle. The Pc pressure fed back to the variable pump 2 is the differential pressure between the pump pressure inside the control valve 22 and the maximum load pressure, and minimizes the effect of pressure loss between the variable pump 2 and the control valve 22 at low temperatures. (For example, refer to Patent Document 1 and FIG. 2).

JP 10-89304 A

However, the hydraulic drive device described in Patent Document 1 has problems in the following points.
(1) Operability The differential pressure reducing valve uses the discharge pressure (P) and maximum load pressure (PLmax) of the variable pump as inputs, and P-PLmax as the output Pc pressure. This Pc pressure is fed back to the pressure compensation valve, and the pressure compensation valve moves to a pressure balancing position.
In the conventional system, a slight change in the pump pressure and the maximum load pressure is detected as a change in the Pc pressure, and the change in the Pc pressure is transmitted to the pressure compensation valve as a signal. There is a time delay for the Pc pressure to change and propagate to the pressure compensation valve with respect to the fluctuation. For this reason, when the flow rate to the actuator is controlled in accordance with the operation of the operator, an excessive flow rate flows instantaneously, or conversely, the flow rate becomes insufficient. In the actual machine, it becomes a shock when the machine is operating. For example, when the operator is operating on a real machine like a hydraulic excavator, there is a problem that the operator feels uncomfortable and the operation itself is difficult to perform, which increases the burden on the operator.

(2) Shuttle valve In patent document 1, in order to detect Pc pressure, it is necessary to take out the highest load pressure from the load pressure of a some actuator. In this case, a shuttle valve is used as a means for taking out the maximum load pressure. Furthermore, since the shuttle valve is only for the purpose of taking out the maximum load pressure and does not require a flow rate, the size is reduced in order to make it as compact as possible. Therefore, since the dimensions are small, it is difficult to process. In addition, n-1 shuttle valves are required for n actuators. In addition, a space for assembling the shuttle valve in the control valve body is required. Furthermore, the thickness of the main body that enables drilling to communicate the pressure signal between the shuttle valves is also necessary. For this reason, the main body of the control valve becomes larger for the space in which the shuttle valve is built. As described above, there is a problem that the cost is considerably increased only with the shuttle valve.

(3) Differential pressure reducing valve In the LS system of Patent Document 1, a differential pressure reducing valve using the discharge pressure of the variable pump and the maximum load pressure as inputs and P-PLmax as the output Pc pressure is used. The pressure reducing valve has a problem in that the structure is complicated, the number of parts is large, and the cost is high compared to a pressure reducing valve with a constant secondary pressure.
The present invention has been made in view of the conventional problems, and an object of the present invention is to provide a hydraulic drive device having an LS system having functions equivalent to those of the prior art and improved operability.
Furthermore, another object of the present invention is to provide a hydraulic drive device having a compact LS system at a low cost while reducing the number of components while improving operability.

In order to solve the above problems, the invention according to claim 1
A variable pump section;
A plurality of actuators driven by the discharge oil of the variable pump unit;
A plurality of directional control valves having a flow rate adjusting function capable of controlling the pressure oil flowing into each of the actuators;
A plurality of pressure compensating valves for compensating the pressure of each of the directional control valves;
A throttle valve provided between the pressure compensation valve and the pressure reducing valve;
It is provided with.
According to the present invention, since the compensation differential pressure to the pressure compensation valve is controlled by the operation of the pressure compensation valve itself, the pressure compensation valve is controlled against fluctuations in the variable pump discharge line pressure and the upstream pressure of the direction switching valve. Responsiveness increases.
Therefore, the shock in the actual machine due to the pressure change of the variable pump discharge line during the operation of the actual machine and the response delay of the pressure compensation valve due to the change in the maximum load pressure is reduced. The operator's discomfort is also reduced, and the actual machine shake due to the shock is reduced, making operation easier.

According to a second aspect of the present invention, the pressure compensation valve includes a valve main body, a spool slidably inserted in an inner hole formed in the valve main body, and a piston that slides on an inner diameter portion of the spool. The spool includes a second pressure receiving area in which a pressure between the throttle valve and the pressure compensating valve acts in the opening direction of the pressure compensating valve, a third pressure receiving area in which the load pressure of the actuator acts, and the throttle valve and the pressure A throttle portion having a function of communicating and shutting off the pressure between the compensation valves with respect to the drain line, and the piston has a first pressure receiving area where the upstream pressure of the direction switching valve acts in the closing direction of the pressure compensation valve. If provided, it is possible to control the discharge pressure of the variable pump to be higher than the maximum load pressure by a constant pressure. Further, even if the discharge flow rate of the variable pump is insufficient with respect to the required flow rate of the control valve, it may have an anti-saturation function that is distributed according to the opening area of the direction switching valve regardless of the magnitude of the load pressure.
According to a third aspect of the present invention, when the throttle section has a shape in which an opening cross-sectional area where the pressure between the throttle valve and the pressure compensation valve communicates with and shuts off the drain line gradually increases, the pressure change at the time of port switching Is performed smoothly, and a shock can be avoided when the pressure compensation valve is switched.

In the present invention, since the pressure compensation valve itself operates by controlling the compensation differential pressure to the pressure compensation valve in the control valve, the responsiveness of the pressure compensation valve to the fluctuations in the discharge pressure and the load pressure is enhanced.
Therefore, the shock in the actual machine due to the pressure change of the variable pump discharge line during the operation of the actual machine and the response delay of the pressure compensation valve due to the change in the maximum load pressure is reduced.
Furthermore, the operator's discomfort is reduced, and the actual machine shake due to the shock is reduced, making it easier to operate.
Even if there is no shuttle valve, and there is no differential pressure reducing valve that detects the differential pump discharge pressure, that is, the differential pump discharge line pressure and the maximum load pressure, it is variable as in the prior art. Control is possible so that the discharge pressure of the pump is higher than the maximum load pressure by a fixed pressure. Further, even when the discharge flow rate of the variable pump is insufficient with respect to the required flow rate of the control valve, it has an anti-saturation function that is distributed according to the opening area of the direction switching valve regardless of the magnitude of the load pressure.
Furthermore, a complicated differential pressure reducing valve is not required because of the large number of parts as in the prior art. By eliminating the shuttle valve and differential pressure reducing valve, it can be manufactured at a lower cost than before, and by eliminating the shuttle valve, the weight of the main body can be reduced.

DESCRIPTION OF EMBODIMENTS Preferred embodiments of a hydraulic drive apparatus according to an embodiment of the present invention will be described in detail below with reference to the accompanying drawings. FIG. 1 is a hydraulic circuit diagram of a hydraulic drive apparatus 10 according to the first embodiment of the present invention. The variable pump unit 11 includes a variable displacement hydraulic pump 13 (hereinafter referred to as a pump) driven by a prime mover 12 such as an engine, a volume changing unit 14 of the pump 13, and discharge oil of the pump 13 to the volume changing unit 14. A pump flow rate adjusting valve 15 that switches the communication between the tank 36 and the volume changing means 14 is provided.
The control valve 17 located downstream of the variable pump unit 11 includes a direction switching valve 18 having a flow rate adjusting function capable of controlling the pressure oil flowing into each of a plurality of actuators 24 and 25 (only two of which are shown). 19, a pressure reducing valve 20, a throttle valve 21, and pressure compensating valves 22 and 23 for compensating pressure of the direction switching valves 18 and 19.

The pump flow rate adjusting valve 15 has a pressure (hereinafter referred to as Pr ′ pressure) between the throttle valve 21 and the pressure compensating valves 22 and 23 in a direction in which pressure oil is supplied to the volume changing means 14, that is, in a direction to reduce the discharge capacity of the pump 13. When the elastic force F of the spring member 38 of the pump flow rate adjusting valve 15 acts, the pressure reducing valve 20 is set in a direction in which the pressure oil in the volume changing means 14 is communicated with the tank 36, that is, in a direction in which the discharge capacity of the pump 13 is increased. Set pressure (hereinafter referred to as Pr pressure) is acting.
Here, the relationship between the elastic force F of the spring member 38 and the pressure receiving chamber area of the pump flow rate adjusting valve 15 with A11 can be replaced by F / A11 = Psp.
Therefore, in FIG. 1, the discharge capacity of the pump 13 is controlled so that Pr = Pr ′ + Psp, that is, Pr ′ = Pr−Psp.

  Pressure compensation valves 22 and 23 are connected in parallel to the discharge oil passages 26 and 27 of the pump 13 via oil passages 28 and 29, and check valves 32 and 23 are respectively connected to output oil passages 30 and 31 of the pressure compensation valves 22 and 23. The pressure oil flowing into the actuators 24 and 25 can be controlled and connected to the direction switching valves 18 and 19 for adjusting the flow rate. The direction switching valves 18 and 19 are connected to the direction switching valves 18 and 19 through 33. Are connected to the actuators 24 and 25, respectively, so that the return oil from the actuators 24 and 25 is returned to the tank 36 from the return oil passages 34 and 35 via the direction switching valves 18 and 19 respectively. Yes.

The control valve 17 is capable of controlling the pressure oil flowing into the actuators 24 and 25 to control the flow rate, the direction switching valves 18 and 19, the pressure reducing valve 20 for reducing the secondary pressure to a constant pressure, the throttle valve 21, Pressure compensation valves 22 and 23 for controlling the differential pressure across the direction switching valves 18 and 19. In this case, the throttle valve 21 is disposed between the pressure reducing valve 20 and the pressure compensating valves 22 and 23. The pressure compensation valves 22 and 23 are located upstream of the direction switching valves 18 and 19, and the pressure upstream of the direction switching valves 18 and 19 (hereinafter referred to as Pin 1 and Pin 2 pressures) in the closing direction of the pressure compensation valves 22 and 23. In the opening direction, Pr ′ pressure and actuator load pressure (hereinafter referred to as PL1 and PL2 pressures) act.
Therefore, the pressure compensation valve 22 is Pin1 = PL1 + Pr ′.
The pressure compensation valve 23 is Pin2 = PL2 + Pr ′.
Therefore, control is performed so that Pr ′ = Pin1-PL1 and Pr ′ = Pin2-PL2 are satisfied.

  Further, in order to obtain an anti-saturation function, the conventional hydraulic drive device is provided with a pump flow rate control valve 38 of the pump 2 and a differential pressure reducing valve 31 that outputs a signal pressure to the pressure compensating valves 4 and 14, Maximum load pressure was led. In order to detect the maximum load pressure in the control valve 22, the shuttle valve 13 was also installed. However, in the hydraulic drive device 10 according to the embodiment of the present invention, the above-described system configuration makes it possible to have an anti-saturation function without arranging a shuttle valve.

FIG. 5 is a longitudinal sectional view of the pressure compensating valves 22 and 23, and FIG. 6 is a detailed view of a portion X in FIG. Since the pressure compensation valves 22 and 23 have the same structure, the pressure compensation valve 22 will be described.
The pressure compensation valve 22 includes a valve main body 60, a spool 62 slidably fitted in an inner hole 61 formed in the valve main body 60, a piston 63 sliding on an inner diameter portion of the spool 62, Is provided. In the closing direction of the pressure compensation valve 22, the upstream pressure Pin1 pressure of the direction switching valve 18 acts on the first pressure receiving area A21. In the opening direction of the pressure compensation valve 22, the load pressure PL1 pressure of the actuator 24 acts on the third pressure receiving area A23. Further, the Pr ′ pressure acts on the second pressure receiving area A22.

Further, a drain line 37 is disposed between the pressure receiving chamber for the PL1 pressure and the pressure receiving chamber for the Pr ′ pressure. As a result, the Pr ′ pressure and the drain line 37 are communicated with or shut off via the pair of throttle portions 64 (see FIG. 6) provided on the spool 62. The throttle part 64 may have a tapered shape as shown in FIG.
As shown in FIGS. 7 and 8, the throttle portion 64 includes a groove portion 65 on the drain line 37 side and a groove 66 connected to the groove portion 65. In this case, the groove part 65 close to the drain line 37 side is wider than the width of the groove part 66, and the depths of these groove parts 65, 66 are uniform. As described above, the throttle portion 64 has a shape in which the opening cross-sectional area where the Pr ′ pressure and the drain line 37 communicate gradually increases as the opening of the variable pump discharge line P pressure and the Pin 1 pressure increases in the spool 62.
Furthermore, as shown in FIG. 9, it is good also as forming the groove parts 67a-67c which are uniform in width and become shallow in order from the drain line 37 side. Further, as shown in FIG. 10, the narrowed portion 64 may be notched groove portions 68a and 68b. As described above, the aperture portion 64 may of course have a shape other than that shown in FIGS. 6 to 10 as long as the opening cross-sectional area gradually increases. Therefore, the pressure compensation valve 22 smoothly changes the Pr ′ pressure when the port is switched.

  In the pump flow rate adjusting valve 15, the Pr ′ pressure acts on the first pressure receiving area A 11 in the direction of opening the cylinder chamber pressure of the volume changing means 14 to the tank 36, and the elastic force of the spring member 38 is also combined. Works. Further, the Pr pressure acts on the second pressure receiving area A12 in the direction in which the pressure is supplied to the cylinder chamber of the volume changing means 14.

The hydraulic drive apparatus 10 according to the embodiment of the present invention is basically configured as described above. Next, the operation and action thereof will be described.
(1) When the required flow rate of the control valve 17 is zero When the direction switching valves 18 and 19 are in the neutral position as shown in FIG. 1, the pressure compensation valves 22 and 23 are in the rightmost position as shown in FIG. The P pressure port in the control valve 17 and the IN side pressure Pin 1 and Pin 2 ports of the direction switching valves 18 and 19 are blocked by the pressure compensating valves 22 and 23. The Pr ′ pressure port and the drain line 37 are also shut off by the pressure compensation valves 22 and 23.
At this time, since there is no flow between the Pr pressure and the Pr ′ pressure, there is no pressure loss, and in this case, Pr ′ = Pr.
At this time, the balance of the force of the pump flow rate adjusting valve 15 is Pr × A12 <(Pr ′ × A11) + F (1) so as to reduce the discharge flow rate of the pump 13.
Thus, the pump flow rate adjusting valve 15 is switched to the right position in FIG. 1, and the volume changing means 14 minimizes the discharge capacity of the pump 13.

(2) When the required flow rate of the control valve 17 increases within the maximum discharge flow rate range of the pump 13 For example, from the state (1), the direction switching valve 18 switches to the left position in FIG. Try to supply. In this case, the load pressure of the actuator 24 (hereinafter referred to as PL1 pressure) acts on the pressure compensation valve 22, and the balance of the force is Pin1 × A21 <(PL1 × A23) + (Pr ′ × A22). (2)
Thus, the pressure compensation valve 22 tries to move to the left position, but the Pr ′ pressure and the drain line 37 communicate with each other through the throttle portion 64a of the pressure compensation valve 22, and a flow is generated in the throttle valve 21. A differential pressure is generated between the Pr pressure and the Pr ′ pressure. Therefore, Equation (2) is expressed as Pin1 × A21 = (PL1 × A23) + (Pr ′ × A22) (2) ′.
(A21 = A22 = A23)
Pin1-PL1 = Pr '(3)
The pressure compensation valve 22 stops at a position that balances.

The balance of the force of the pump flow rate adjusting valve 15 of the pump 13 is Pr × A12> (Pr ′ × A11) + F (4)
Thus, the pump flow rate adjusting valve 15 is switched to the position shown in FIG. 1, and the volume changing unit 14 controls to increase the discharge flow rate of the pump 13 so as to satisfy the following expression (5).
Pr × A12 = (Pr ′ × A11) + F (A12 = A11)
Pr−Psp = Pr ′ (5)

(3) When the required flow rate of the control valve 17 decreases within the maximum discharge flow rate range of the pump 13 For example, when the opening degree of the direction switching valve 18 is reduced from the state of (2), the pressure loss of the direction switching valve 18 is reduced. The Pin1 pressure increases momentarily, and the balance of the force of the pressure compensation valve 22 is Pin1 × A21> (PL1 × A23) + (Pr ′ × A22) (6)
It becomes.
Therefore, the pressure compensation valve 22 moves in a direction to reduce the opening degree that communicates the P pressure port and the Pin1 pressure port, and the area of the throttle portion 64a of the pressure compensation valve 22 that communicates the Pr ′ pressure and the drain line 37. Becomes smaller and the drain amount itself becomes smaller. In this case, the pressure compensation valve 22 moves from the left position in FIG. 1 to an intermediate position close to the center position.
Therefore, the Pr ′ pressure is increased, and the pressure compensation valve 22 stops at a balanced position so that the following expression (7) is established.
Pin1 × A21 = (PL1 × A23) + (Pr ′ × A22)
Pin1-PL1 = Pr ′ (7) (same as (8))
At the same time, the balance of the force of the pump flow control valve 15 of the pump 13 is Pr × A12 <(Pr ′ × A11) + F (8)
Thus, the pump flow rate adjusting valve 15 moves to the right position in FIG. 1, and the volume changing means 14 controls the discharge flow rate of the pump 13 to decrease so as to satisfy the following expression (9).
Pr × A12 = (Pr ′ × A11) + F
Pr−Psp = Pr ′ (9) (same as equation (5))
The hydraulic drive device 10 according to the present embodiment is configured so that the differential pressure across the directional control valves 18 and 19 is changed to a target compensation pressure (Pr−Psp) as in the above formulas (3), (5), (7), and (9). Control to become.
In the single operation as described above, PL1 pressure = maximum load pressure PLmax,
Since Pin1 pressure = pump pressure P, P−PLmax = Pr ′ = Pr−Psp.
In this case, Pr ′ means a differential pressure between the pump pressure P and the maximum load pressure PLmax.

(4) When the required flow rate of the control valve 17 is greater than or equal to the maximum discharge flow rate of the pump 13 Here, the actuators 24 and 25 are operated simultaneously, and the saturation state is reached where the maximum discharge flow rate of the pump 13 is lower than the required flow rate of the control valve 17. Explain when.
At this time, the load pressure (hereinafter referred to as PL2 pressure) of the actuator 25 is set larger than the PL1 pressure.
The balance of pressure acting on the pressure compensation valve 22 of the actuator 24 is
From Pin1 = PL1 + Pr ′ Pin1-PL1 = Pr ′ (10)
The balance of pressure acting on the pressure compensation valve 23 of the actuator 25 is
From Pin2 = PL2 + Pr 'Pin2-PL2 = Pr' (11)

The pressure compensation valves 22 and 23 move to the positions on the left side of FIG. 1 because the flow rate is insufficient and the pressures of Pin 1 and Pin 2 do not increase by Pr ′ pressure with respect to the PL 1 pressure and PL 2 pressure. In both the pressure compensation valves 22 and 23, the Pr 'pressure communicates with the drain line 37, so that the Pr' pressure is lower than in the cases (2) and (3).
Thereby, the balance of the force of the pump flow rate adjusting valve 15 of the pump 13 is Pr × A12> (Pr ′ × A11) + F (12)
Thus, the pump flow rate adjusting valve 15 is switched to the left position in FIG. 1 and the volume changing means 14 is controlled to increase the discharge flow rate of the pump 13. Since the required flow rate is large, the equation (12) is Pr-Psp> Pr ′ (13)
The relationship remains. Therefore, the pump 13 continues to discharge the maximum discharge flow rate.

Further, PL2> PL2 and PL2 pressure = PLmax.
Therefore, Pin 2 pressure of the pressure compensation valve 23 on the high load side = PLmax + Pr ′ (14)
At this time, Pin1 pressure of the pressure compensation valve 22 on the low load pressure side = PL1 + Pr ′ (15)
And Pin2 pressure> Pin1 pressure.
Therefore, P pressure = Pin 2 pressure of the hydraulic drive device 10 of the present invention (16)
The pressure compensation valve 23 on the high load side has the opening cross-sectional area at the fully open position so as to satisfy the expressions (14) and (16).
Further, the pressure compensation valve 22 on the low pressure side stops in a balanced manner at a position where the P pressure is reduced to the Pin 1 pressure (opening degree is reduced).
In the hydraulic drive device 10 of the present invention, when a plurality of actuators 24 and 25 are simultaneously operated and the discharge flow rate of the pump 13 is insufficient with respect to the required flow rate of the control valve 17, the Pr 'pressure is the target compensation differential pressure Pr-Psp. The pressure is reduced to a pressure lower than the pressure compensation pressures of the pressure compensation valves 22 and 23. Therefore, as in the prior art, even if the discharge flow rate of the pump 13 is insufficient, it has an anti-saturation function in which the flow rate is distributed according to the opening area of the direction switching valves 18 and 19 regardless of the load.

In the conventional system, the fluctuation of the pump pressure and the maximum load pressure is detected via the differential pressure reducing valve, so that the fluctuation of the Pc pressure to the pressure compensation valve is delayed. Therefore, it is desired to increase or decrease the flow rate of the actuator according to the operation of the operator. However, in actuality, the change in the flow rate to the actuator is instantaneously delayed, the surplus flow rate or the flow rate becomes insufficient, the actuator speed changes suddenly, and a shock occurs.
In a work machine such as a hydraulic excavator, multiple actuators with different loads can be operated at the same time, or from simultaneous operation to any single actuator operation, or conversely, from single operation to simultaneous operation of multiple actuators There is frequent work to do. Therefore, even if the load pressure of an arbitrary actuator is almost constant, the pump pressure frequently fluctuates due to the load pressure of other actuators. Then, even if the operator intends to operate the actuator at a constant speed, a shock is generated in the operation state of other actuators, and the operator himself feels uncomfortable. On the other hand, it may move, and it is necessary to be careful during operation, which increases the burden.

In the hydraulic drive device 10 of the present invention, the compensation differential pressure of the pressure compensation valves 22 and 23 is the Pr ′ pressure. For example, when the operation of the actuator 25 is stopped from the state in which the actuators 24 and 25 in FIG. 1 are operated at a constant speed, the pressure compensation valve 22 is in a state where the flow rate is stably controlled (left end position in FIG. 1). Since the P pressure suddenly increases, the intermediate position moves to the right end position. At this time, while moving, the Pr ′ pressure shifts from a state communicating with the drain line 37 to a blocking state, but the Pr ′ pressure is controlled by the operation of the pressure compensation valves 22 and 23 themselves. Response delay with respect to fluctuations in P pressure and maximum load pressure is smaller than in conventional systems. Therefore, the excessive flow volume to the actuator 24 can be suppressed, and the shock of the actual machine can be suppressed as compared with the conventional system.
Here, in the pressure compensation valves 22 and 23, a pressure receiving chamber of Pr ′ pressure and a pressure receiving chamber of PL1 pressure,
A structure in which the pressure receiving chamber for Pr ′ pressure and the pressure receiving chamber for PL2 pressure are installed in reverse may be used. Further, the communication timing of the aperture portions 64a and 64b, the gain of the opening area, and the maximum opening area can be arbitrarily set.

FIG. 2 is a hydraulic circuit diagram of the hydraulic drive device 70 according to the second embodiment of the present invention. In FIG. 2, the same components as those of FIG. Omitted. The same shall apply hereinafter.
In the hydraulic circuit diagram of the hydraulic drive device 70, the primary pressure of the pressure reducing valve 20 is not the P pressure of the control valve 17, but the discharge pressure of the fixed pump 71 driven by the prime mover 12 such as the engine together with the pump 13. It is characterized by being. The relief valve 72 is used for the fixed pump 71.
FIG. 3 is a hydraulic circuit diagram of the hydraulic drive device 80 according to the third embodiment of the present invention.
In this hydraulic circuit diagram, the spring member 82 of the pump flow rate adjusting valve 81 is caused to act in a direction in which the pressure of the volume changing means 14 communicates with the tank 36, and the throttle valve 21 and the pressure compensation are provided on the other side of the pump flow rate adjusting valve 81. A pressure between the valves 22 and 23 (hereinafter referred to as Pr ′ pressure) may be introduced. At this time, if the elastic force of the spring member 82 is equivalent to the differential pressure Pr−Psp between the Pr pressure of FIG. 1 and the elastic force F of the spring member 38 (F / A11 = Psp), it is equivalent to FIG. It becomes. At this time, the oil chamber in which the spring member 82 is applied has a drain pressure.

FIG. 4 is a hydraulic circuit diagram of a hydraulic drive device 90 according to the fourth embodiment of the present invention. In this hydraulic circuit diagram, the pressure reducing valve 20 (see FIG. 1) may be eliminated, and the Pr pressure may be set by the engine speed detection valve 91. The engine speed detection valve 91 includes a throttle valve 92 through which the discharge flow rate of the fixed pump 71 passes, and a differential pressure reducing valve 93 that detects a differential pressure across the throttle valve 92. The differential pressure before and after the throttle valve 92 detected by the differential pressure reducing valve 93 is used as the Pr pressure of the target compensation differential pressure Pr-Psp. Since the discharge flow rate of the fixed pump 71 changes in conjunction with the rotational speed of the prime mover 12 such as the engine, the differential pressure across the throttle valve 92 also changes in conjunction with the prime mover 12 such as the engine. Therefore, in this hydraulic drive device 90, a system in which the target compensation differential pressure also changes corresponding to the change in the flow rate of the pump 13 according to the rotational speed of the prime mover 12 such as an engine.

According to the present invention, since the Pr 'pressure, which is the differential pressure between the variable pump discharge line pressure P and the maximum load pressure PLmax in the control valve, is controlled, the pressure compensation valve against fluctuations in the P pressure and the load pressures Pin1 and Pin2 Responsiveness increases.
Therefore, the shock in the actual machine due to the pressure change of the variable pump discharge line during the operation of the actual machine and the response delay of the pressure compensation valve due to the change in the maximum load pressure is reduced. The operator's discomfort is also reduced, and the actual machine shake due to the shock is reduced, making operation easier.
It can be set as the structure which does not need the shuttle valve which is a means to detect the maximum load pressure PLmax like a prior art.

Even in a configuration without a shuttle valve and without a differential pressure reducing valve for detecting the differential pump discharge pressure, that is, the differential pressure between the variable pump discharge line pressure P and the maximum load pressure PLmax, it is variable as in the prior art. Control can be performed such that the pump discharge pressure P is higher than the maximum load pressure PLmax by a fixed pressure. Further, even when the discharge flow rate of the variable pump is insufficient with respect to the required flow rate of the control valve, it has an anti-saturation function that is distributed according to the opening area of the direction switching valve regardless of the magnitude of the load pressure.
Furthermore, a complicated differential pressure reducing valve is not required because of the large number of parts as in the prior art. The abolition of the shuttle valve and the differential pressure reducing valve enables manufacturing at a lower cost than the conventional one, and the weight of the main body can be reduced by eliminating the shuttle valve.

1 is a hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention. FIG. 5 is a hydraulic circuit diagram of a hydraulic drive device according to a second embodiment of the present invention. FIG. 5 is a hydraulic circuit diagram of a hydraulic drive device according to a third embodiment of the present invention. FIG. 6 is a hydraulic circuit diagram of a hydraulic drive device according to a fourth embodiment of the present invention. It is a longitudinal cross-sectional view which shows schematic structure of the pressure compensation valve of FIG. FIG. 6 is an enlarged detail view of a main part of the spool of FIG. 5. FIG. 6 is an enlarged detail view of a main part of another spool of FIG. 5. It is a principal part perspective view of the groove part of FIG. It is a principal part perspective view of the other groove part of FIG. It is a principal part perspective view of the other groove part of FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10, 70, 80, 90 Hydraulic drive device 11 Variable pump part 12 Motor | operator 13 Variable displacement hydraulic pump 14 Volume change means 15, 81 Pump flow rate adjustment valve 17 Control valve 18, 19 Directional switching valve 20 Pressure reducing valve 21, 92 Throttle valve 22, 23 Pressure compensation valve 24, 25 Actuator 60 Valve body 62 Spool 63 Piston 64 Restriction part 71 Fixed pump 91 Engine speed detection valve 93 Differential pressure reducing valve

Claims (3)

A variable pump section;
A plurality of actuators driven by the discharge oil of the variable pump unit;
A plurality of directional control valves having a flow rate adjusting function capable of controlling the pressure oil flowing into each of the actuators;
A plurality of pressure compensating valves for compensating the pressure of each of the directional control valves;
A throttle valve provided between the pressure compensation valve and the pressure reducing valve;
A hydraulic drive device comprising: The hydraulic drive device according to claim 1, wherein
The pressure compensation valve includes a valve body, a spool that is slidably inserted into an inner hole formed in the valve body, and a piston that slides on an inner diameter portion of the spool.
The spool includes a second pressure receiving area in which a pressure between the throttle valve and the pressure compensating valve acts in the opening direction of the pressure compensating valve, a third pressure receiving area in which the load pressure of the actuator acts, and the throttle valve and the pressure compensating valve And a throttle part having a function of communicating and blocking between the drain line and the drain line,
The hydraulic drive apparatus according to claim 1, wherein the piston is provided with a first pressure receiving area in which the upstream pressure of the direction switching valve acts in a closing direction of the pressure compensation valve. In the hydraulic drive unit according to claim 1 or 2,
The hydraulic drive device according to claim 1, wherein the throttle portion has a shape in which an opening cross-sectional area where the pressure between the throttle valve and the pressure compensation valve communicates with or shuts off the drain line gradually increases.

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