JP2009014122A - Hydraulic drive of construction machine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To enhance energy efficiency by surely preventing cavitation from being generated even if the discharge flow rate of a hydraulic pump is low, such as, for example, when deceleration of a heavy load inertia body is operated and reducing useless pressure loss if discharge the flow rate of the hydraulic pump is high, such as, for example, when a high flow rate actuator is driven in a hydraulic drive of a construction machinery. <P>SOLUTION: The hydraulic drive is provided with a refill circuit 38 in a pair of actuator oil pathways 37a, 37b of a swiveling motor 12 in a control valve 6. A tank return circuit 7 of the control valve 6 includes a return line 42 which makes returned oil from the control valve 6 reflux to a tank, an oil cooler 43 and a back pressure generator 44 which are provided in the return line 42 and is connected with the refill circuit 38 on the upstream side of the back pressure generator 44. The back pressure generator 44 is configured as a pressure control valve which keeps the differential pressure between pressures of the most upstream and most downstream side of the return line 42 constant. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a hydraulic drive device used in construction machinery such as a hydraulic excavator, and in particular, a back pressure generating device is provided in a tank return circuit of a control valve, and pressure oil is supplied during inertial driving of an actuator, such as during turning deceleration of a turning motor. The present invention relates to a hydraulic drive device for a construction machine that enables prompt replenishment and prevents the occurrence of cavitation.

  There exists a thing described in patent document 1 as this kind of hydraulic drive device. The hydraulic drive device is supplied to an engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by pressure oil discharged from the hydraulic pump, and a plurality of actuators from the hydraulic pump. A control valve having a plurality of flow control valves for controlling the flow rate of pressure oil, a replenishment circuit provided in an actuator oil passage of at least one of the plurality of actuators, and a return for circulating the return oil from the control valve to the tank And a tank return circuit having an oil cooler and a back pressure generating device provided in the return line and a back pressure generating device, to which an upstream side replenishment circuit is connected.

  In addition, hydraulic drive devices for construction machines such as hydraulic excavators generally include pump control means for controlling the capacity of a hydraulic pump in accordance with the required flow rates of a plurality of flow control valves. As this pump control means, for example, a patent As described in Document 2, a load sensing system is known that controls the capacity of a hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of a plurality of actuators by a target differential pressure.

JP 2005-291312 A Japanese Patent Laid-Open No. 10-196604

  However, there are the following problems in the above-described prior art.

  In the case of a hydraulic drive device for a construction machine, the actuator is often driven by a heavy inertial body, and cavitation is likely to occur during a deceleration operation or a stop operation of such a heavy inertial body. For example, in the upper swing body of a hydraulic excavator, during a deceleration operation, the flow control valve for swing is returned to the neutral position side to throttle the meter-in oil passage to reduce the amount of oil supplied. Since the swing motor tends to continue to rotate at the same speed as before, the pressure of the actuator oil passage on the pressure oil supply side between the flow control valve and the swing motor decreases, and cavitation occurs. Cavitation occurs in the same manner when the upper swing body is stopped, when other heavy inertia bodies are decelerated, and when the stop operation is performed.

  Normally, an actuator oil passage of an actuator that drives a heavy object inertial body such as a swing motor is provided with a replenishment circuit. When the pressure of the actuator oil passage on the pressure oil supply side drops, the tank pressure is supplied via the replenishment circuit. Oil is replenished to prevent cavitation.

  Here, since the discharge flow rate of the hydraulic pump is controlled according to the required flow rate of the flow control valve, the discharge flow rate of the hydraulic pump decreases during the deceleration operation or the stop operation, and the flow rate circulating to the tank also decreases. As a result, the back pressure of the tank return oil passage of the control valve is reduced, so if there is only a replenishment circuit, the pressure oil cannot be replenished quickly when the pressure of the actuator oil passage on the pressure oil supply side drops, and cavitation The prevention effect is reduced. When a back pressure valve is provided in the tank return circuit as described in Patent Document 1, a sufficient back pressure is secured by the back pressure valve even when the pump discharge flow rate is reduced and the flow rate circulating to the tank is reduced. Replenishment is reliably performed by the pressure, and the occurrence of cavitation can be reliably prevented.

  Further, as described in Patent Document 2, when the pump control means is a load sensing system, the meter-in oil passage of the flow control valve is throttled and the discharge pressure of the hydraulic pump increases even during the deceleration operation. The maximum load pressure decreases as the pressure of the actuator oil passage on the pressure oil supply side of the swing motor decreases, and the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure increases. Decreases to the minimum. Therefore, in the load sensing method, the occurrence of cavitation is particularly prominent, and it is useful to provide a back pressure valve as described in Patent Document 1 in the tank return circuit. Even when the flow rate flowing back to the tank is reduced to the minimum, the back pressure valve ensures the replenishment and the occurrence of cavitation is reliably prevented.

By the way, when the conventional tank return circuit as described in Patent Document 1 is provided, the return oil from the control valve passes through the back pressure valve and returns to the tank via the oil cooler. When the pressure loss generated by the back pressure valve is ΔP ₁ , the pressure loss generated by the oil cooler is ΔP ₂ , and the pressure loss generated in the entire tank return circuit is ΔPc, the relationship ΔPc = ΔP ₁ + ΔP ₂ holds. Here, ΔP ₁ is substantially constant regardless of the flow rate Q, but ΔP ₂ increases as the flow rate Q increases, and ΔPc has a characteristic obtained by adding ΔP ₁ and ΔP ₂ . Here, ΔP ₁ is a value that allows quick replenishment during deceleration operation, that is, when the discharge flow rate of the hydraulic pump is minimized, so that unpleasant cavitation does not occur. As a result, the pressure loss ΔPc of the entire tank return circuit is almost the same value as the pressure Ps necessary for preventing cavitation when the flow rate Q is small, but in the case of driving a large flow rate actuator, When Q is large, it becomes larger than necessary due to the effect of ΔP ₂ that increases as the flow rate Q increases. This pressure loss that is larger than necessary is a wasteful energy loss, resulting in a deterioration in the energy efficiency of the hydraulic drive device.

  The object of the present invention is to reliably prevent cavitation even when the discharge flow rate of the hydraulic pump is small, such as during deceleration operation of a heavy object inertial body, and the discharge flow rate of the hydraulic pump such as when driving a large flow rate actuator. It is an object of the present invention to provide a hydraulic drive device for a construction machine that can reduce useless pressure loss in a large case and improve energy efficiency.

  (1) To achieve the above object, the present invention includes an engine, a variable displacement hydraulic pump driven by the engine, and a plurality of actuators driven by pressure oil discharged from the hydraulic pump, A control valve having a plurality of flow control valves that control flow rates of pressure oil supplied from the hydraulic pump to a plurality of actuators, and a pump that controls the capacity of the hydraulic pump according to the required flow rates of the plurality of flow control valves In the hydraulic drive apparatus for a construction machine, comprising a control means, a supply circuit provided in an actuator oil passage of at least one of the plurality of actuators, a return line for circulating the return oil from the control valve to the tank, the return An oil cooler and a back pressure generator provided in a line, and the upstream side of the back pressure generator A tank return circuit to which a supply circuit is connected, and the back pressure generating device is a pressure control valve that controls the differential pressure between the most upstream pressure and the most downstream pressure in the return line to be constant. And

  By providing the back pressure generator in this way, it is possible to reliably prevent cavitation even when the discharge flow rate of the hydraulic pump is small, such as during the deceleration operation of the heavy object inertial body. In addition, the back pressure generator is equipped with a pressure control valve that controls the differential pressure between the most upstream pressure and the most downstream pressure in the return line so that the hydraulic pump discharges when driving a large flow actuator. Even when the flow rate is large and the tank return flow rate is large, the pressure loss is kept constant, so that wasteful pressure loss can be reduced and energy efficiency can be improved.

  (2) In order to achieve the above object, the present invention provides an engine, a variable displacement hydraulic pump driven by the engine, and a plurality of actuators driven by pressure oil discharged from the hydraulic pump. A control valve having a plurality of flow rate control valves for controlling the flow rate of pressure oil supplied from the hydraulic pump to the plurality of actuators, and a discharge pressure of the hydraulic pump being a target differential pressure from a maximum load pressure of the plurality of actuators In a hydraulic drive device for a construction machine comprising a load sensing type pump control means for controlling the capacity of the hydraulic pump so as to be higher, a replenishment circuit provided in an actuator oil path of at least one of the plurality of actuators; A return line for circulating the return oil from the control valve to the tank; And a tank return circuit connected to the replenishment circuit upstream of the back pressure generator, and the back pressure generator is connected to the return line. It is assumed that the pressure control valve controls the pressure difference between the most upstream pressure and the most downstream pressure to be kept constant.

  In the hydraulic drive apparatus having the load sensing type pump control means for controlling the capacity of the hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure by the target differential pressure by providing the back pressure generating device in this way. Even if the pump discharge flow rate is reduced to the minimum by load sensing control during the deceleration operation of the heavy inertial body and the flow rate circulating to the tank is reduced, the occurrence of cavitation can be reliably prevented. Further, as described in (1) above, by providing a pressure control valve that controls the differential pressure between the most upstream pressure and the most downstream pressure in the return line as a back pressure generating device, When the discharge flow rate of the hydraulic pump is large, such as when the flow rate actuator is driven, wasteful pressure loss can be reduced and energy efficiency can be improved.

  (3) In the above (1) or (2), preferably, the pilot hydraulic power source is provided on the downstream side of the pilot hydraulic power source, and the pressure of the pilot hydraulic power source is adjusted when the gate lock lever is in an operable state. A gate lock valve that outputs to the downstream side and outputs the tank pressure to the downstream side when the gate lock lever is in an inoperable state, and the pressure control valve includes the first pressure receiving chamber and the throttle for the opening direction action. The second and third pressure receiving chambers acting in the direction and the spring means acting in the opening direction, the most upstream pressure of the return line is guided to the first pressure receiving chamber, and the return line is fed to the second pressure receiving chamber. The pressure downstream of the gate lock valve is guided to the third pressure receiving chamber, and the difference between the oil pressure of the third pressure receiving chamber generated by the pressure and the spring force of the spring means By the return Setting the most upstream pressure and the target differential pressure downstream of the pressure down.

  As a result, the pressure control valve controls the pressure difference between the most upstream pressure and the most downstream pressure in the return line to be kept constant.

  When the gate lock lever is set to the inoperable position, the pressure in the third pressure receiving chamber becomes the tank pressure, so that the pressure control valve is forcibly switched to the communication position by the action of the spring means, and the pressure loss in the tank return circuit is reduced to oil. Only the pressure loss generated by the cooler can be suppressed, and the energy efficiency can be further improved.

  (4) Further, in the above (1) or (2), the pressure control valve includes a first pressure receiving chamber acting in the opening direction, a second pressure receiving chamber acting in the throttle direction, and spring means acting in the throttle direction. The most upstream pressure of the return line is guided to the first pressure receiving chamber, the most downstream pressure of the return line is guided to the second pressure receiving chamber, and the most upstream pressure of the return line is A target differential pressure of the most downstream pressure may be set.

  As a result, the pressure control valve controls the pressure difference between the most upstream pressure and the most downstream pressure in the return line to be kept constant.

  According to the present invention, it is possible to reliably prevent cavitation even when the discharge flow rate of the hydraulic pump is small, such as during deceleration operation of a heavy object inertial body, and the discharge flow rate of the hydraulic pump such as when a large flow actuator is driven. When it is large, useless pressure loss can be reduced and energy efficiency can be improved.

  In addition, when the gate lock lever is set to the inoperable position, the pressure control valve is forcibly switched to the communication position, so that the pressure loss of the tank return circuit can be suppressed to only the pressure loss generated by the oil cooler, and further the energy Efficiency can be improved.

Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the following embodiments, the present invention is applied to a hydraulic drive device of a hydraulic excavator.
<Embodiment 1>
~Constitution~
FIG. 1 is a hydraulic circuit diagram showing a configuration of a hydraulic drive device for a construction machine (hydraulic excavator) in the first embodiment (Embodiment 1) of the present invention.

  The hydraulic drive apparatus according to the present embodiment includes an engine 1, a variable displacement main hydraulic pump (main pump) 2 driven by the engine 1, and a fixed displacement driven by the engine 1 in conjunction with the main pump 2. And a plurality of actuators 10, 11, 12 driven by pressure oil discharged from the main pump 2, a control valve 6, and a tank return circuit 7 of the control valve 6.

  The actuators 10, 11, and 12 are, for example, a boom cylinder that drives a boom of a front working machine of a hydraulic excavator, an arm cylinder that drives an arm, and a turning motor that turns an upper turning body. The hydraulic drive device of a hydraulic excavator is usually equipped with other actuators such as a bucket cylinder that drives the bucket of the front working machine, a traveling motor that drives the lower traveling body, and a swing cylinder that swings the front working machine. However, their illustration and description are omitted.

  The control valve 6 is connected to the supply oil passage 2a of the main pump 2, and has a plurality of valve sections 17, 18, 19 for controlling the direction and flow rate of pressure oil supplied from the main pump 2 to each actuator, and a plurality of valve sections. Among the load pressures of the actuators 10, 11, and 12, a plurality of shuttle valves 33 a, 33 b, and 33 c that select the highest load pressure (hereinafter referred to as the maximum load pressure) PLmax and output to the signal oil passage 27, and the main pump 2 A main relief valve 32 provided in the supply oil passage 2a for limiting the maximum discharge pressure (maximum pump pressure) of the main pump 2, and a differential pressure PLS between the discharge pressure (pump pressure) Pd of the main pump 2 and the maximum load pressure PLmax. Is output as an absolute pressure, and the discharge flow rate of the main pump 2 when the differential pressure PLS between the pump pressure Pd and the maximum load pressure PLmax exceeds a certain value. And an unload valve 31 that keeps the differential pressure PLS below a certain value set by the spring 31a. The valve section 17 includes a flow control valve 34a and a pressure compensation valve 35a, the valve section 18 includes a flow control valve 34b and a pressure compensation valve 35b, and the valve section 19 includes a flow control valve 34c and a pressure compensation valve 35c. Has been. The output pressure of the differential pressure reducing valve 30 is led to the valve-opening side pressure receiving parts 36a, 36b, 36c for setting the target compensation differential pressure of the pressure compensating valves 35a, 35b, 35c, and the absolute pressure PLS (discharge pressure of the main pump 2) is introduced. The target compensation differential pressure is set by the differential pressure between Pd and the maximum load pressure PLmax. Outlets of the unload valve 31 and the main relief valve 32 are connected to the tank oil passage 39 in the control valve 6, and the tank oil passage 39 is connected to the tank via the tank return circuit 7.

  The control valve 6 also has a replenishment circuit 38 provided in a pair of actuator oil passages 37 a and 37 b of the swing motor 12. The replenishing circuit 38 replenishes pressure oil when the pressure of the actuator oil passage on the inlet side of the swing motor 12 becomes extremely low during inertial driving such as when the swing motor 12 is decelerated, and the actuator oil passage. Check valves 38a and 38b are provided between 37a and 37b and the tank oil passage 39 in the control valve 6 and permit only the flow of pressure oil from the tank oil passage 39 to the actuator oil passages 37a and 37b. .

  The downstream side of the tank oil passage 39 is connected to the tank return circuit 7, and the return oil from the actuators 10, 11, 12 circulates to the tank via the tank oil passage 39 and the tank return circuit 7. Further, the tank return circuit 7 generates an appropriate back pressure and can be replenished quickly by the replenishment circuit 38. Details of the configuration of the tank return circuit 7 will be described later.

  The hydraulic drive apparatus according to the present embodiment is also connected to a supply oil passage 3a of the pilot pump 3, and outputs an engine speed detection valve device 5 that outputs an absolute pressure Pa according to the discharge flow rate of the pilot pump 3. A pilot hydraulic power source 21 having a pilot relief valve 13 connected to the downstream side of the number detection valve device 5 and keeping the pressure of the first pilot oil passage 20 constant, and connected to a downstream side of the pilot hydraulic pressure source 21, The electromagnetic switching valve 14 that is ON / OFF controlled by the open / close state of the gate lock lever 124 (FIG. 3) provided at the driver's seat entrance and the second pilot oil passage 22 are connected, and the hydraulic pressure of the pilot hydraulic source 21 is reduced to the original pressure. Remote control valves 24, 25, and 26 that generate control pilot pressures a to f for operating the flow control valves 34a, 34b, and 34c, and a main pump And a pump tilt control device 4 for controlling the tilt angle (capacity) of the pump 2.

  When the gate lock lever 124 (FIG. 3) is in an operable state (lowered position), the electromagnetic switching valve 14 cuts off the communication between the second pilot oil passage 22 and the tank, and the first pilot oil passage 20 and the first 2 The pilot hydraulic power source is switched to the first position (ON position) on the left side of the figure where the pilot oil passage 22 is communicated, and is connected to the downstream side of the pilot oil passage 22 or is kept constant by the pilot relief valve 13 at the remote control valves 24, 25, 26. When the pressure of No. 21 is introduced and the gate lock lever 124 (FIG. 3) is in an inoperable state (raised position), the communication between the first pilot oil passage 20 and the second pilot oil passage 22 is cut off, and the second pilot The oil passage 22 is switched to the second position (OFF position) on the right side of the figure for communicating with the tank and connected to the downstream side of the oil path 22 or the remote control valves 24, 25, 26 are communicated with the tank. That. The position of the electromagnetic switching valve 14 is switched by the gate lock lever 124 when, for example, a switch (not shown) is provided between the solenoid of the electromagnetic switching valve 14 and the power source, and the gate lock lever 124 is in an operable state (lowered position). The switch is turned on (closed) to excite the solenoid, and when the gate lock lever 124 is in an inoperable state (up position), the switch is turned off (opened) to release the solenoid.

  The engine speed detection valve device 5 has a flow rate detection valve 5a and a differential pressure reducing valve 5b. The input side of the flow rate detection valve 5a is connected to the supply oil passage 3a of the pilot pump 3, and the output side of the flow rate detection valve 5a. Is connected to the first pilot oil passage 20. The flow rate detection valve 5a converts the flow rate of the pressure oil flowing through the supply oil passage 3a of the pilot pump 3 into the differential pressure across the variable throttle 5c, and the differential pressure reducing valve 5b outputs the differential pressure across the front and back as an absolute pressure Pa. . Since the discharge flow rate of the pilot pump 3 varies depending on the rotational speed of the engine 1, the rotational speed of the engine 1 can be detected by detecting the flow rate (the differential pressure across the variable throttle 50a). Further, the variable throttle portion 5c is configured such that the throttle diameter (area) increases as the front-rear differential pressure increases, and the increase degree of the front-rear differential pressure is moderate.

  The pump tilt control device 4 includes a horsepower control tilt actuator 4a, an LS control valve 4b, and an LS control tilt actuator 4c.

  The horsepower control tilt actuator 4a reduces the tilt angle of the main pump 2 when the discharge pressure of the main pump 2 increases, and limits the input torque of the main pump 2 so as not to exceed a preset maximum torque. This limits the horsepower consumed by the main pump 2 and prevents the engine 1 from being stopped (engine stall) due to overload.

  The LS control valve 4b has pressure receiving portions 4d and 4e facing each other, and the absolute pressure Pa generated by the differential pressure reducing valve 5b of the engine speed detection valve device 5 is the pressure difference 4d. The absolute pressure PLS (the differential pressure between the discharge pressure Pd of the main pump 2 and the maximum load pressure PLmax) generated by the differential pressure reducing valve 30 is guided to the pressure receiving part 4e, and the absolute pressure PLS. Becomes higher than the absolute pressure Pa (PLS> Pa), the pressure of the pilot hydraulic source 21 is guided to the LS control tilt actuator 4c to reduce the tilt angle of the main pump 2, and the absolute pressure PLS is lower than the absolute pressure Pa. Then (PLS <Pa), the LS control tilt actuator 4c is communicated with the tank to increase the tilt angle of the main pump 2, so that the discharge pressure Pd of the main pump 2 is higher than the maximum load pressure PLmax by the target differential pressure Pa. Become Controlling the tilting amount of the main pump 2 (displacement). The control valve 4b and the LS control tilt actuator 4c are arranged so that the discharge pressure Pd of the main pump 2 is higher than the maximum load pressure PLmax of the plurality of actuators 10, 11, 12,. A load sensing type pump control means for controlling the tilt of the pump 2 is configured.

  Here, since the absolute pressure Pa is a value that changes in accordance with the engine speed, the absolute pressure Pa is used as the target differential pressure of the load sensing control, and the target compensated differential pressure of the pressure compensation valves 35a, 35b, and 35c is used as the main pump. By setting the absolute pressure PLS as the differential pressure between the discharge pressure Pd of 2 and the maximum load pressure PLmax, the actuator speed can be controlled according to the engine speed. Further, as described above, the variable throttle portion 5c of the flow rate detection valve 5a of the engine speed detection valve device 5 has its throttle diameter (area) increased as the front-rear differential pressure increases, and the degree of increase in the front-rear differential pressure. Is configured to be moderate. As a result, the saturation phenomenon according to the engine speed can be improved, and good fine operability can be obtained when the engine speed is set low. This point is detailed in Japanese Patent Laid-Open No. 10-196604.

  The set pressure of the spring 31a of the unload valve 31 is the absolute pressure Pa (load sensing) generated by the differential pressure reducing valve 5b of the engine speed detecting valve device 5 when the engine 1 is at the rated maximum speed (for example, 2000 rpm). It is set to be higher than the control target differential pressure.

  The feature of this embodiment is the configuration of the tank return circuit 7 of the control valve 6, which is connected to the tank oil passage 39 of the control valve 6 and the return line 42. The oil cooler 43, a back pressure generator 44 connected in series to the upstream side of the oil cooler 43, and a bypass check valve 45 connected in parallel to the oil cooler 43.

  The return oil from the control valve 6 passes through the back pressure generator 44 and returns to the tank via the oil cooler 43 for cooling the hydraulic oil. The oil cooler 43 is equipped with a bypass check valve 45 in parallel with the oil cooler 43, and is opened when the oil temperature is low, the viscosity of the hydraulic oil is high, and the pressure loss of the oil cooler 43 is high. Yes. Since the pressure loss of the oil cooler 43 is often not so high at the oil temperature normally used, the bypass check valve 45 is closed.

  The back pressure generator 44 is a pressure control valve that controls the pressure difference between the most upstream pressure Pup in the return line 42 (corresponding to the back pressure generated by the back pressure generator 44) and the most downstream pressure Pdown to be constant. The pressure control valve 44 is hereinafter referred to as a pressure control valve 44. The pressure control valve 44 includes a first pressure receiving chamber 44a that operates in the opening direction and a second pressure receiving chamber 44b that operates in the throttle direction and is located on the side facing the first pressure receiving chamber 44a. And a spring 44c (spring means) acting in the throttle direction located on the same side as the second pressure receiving chamber 44b. The most upstream pressure Pup of the return line 42 is guided to the first pressure receiving chamber 44a, and the second The pressure Pdown on the most downstream side of the return line 42 is guided to the pressure receiving chamber 44b, and the target differential pressure between the pressure Pup on the most upstream side of the return line 42 and the pressure Pdown on the most downstream side is set by the spring force of the spring 44c. That is, the spring 44c functions as a target differential pressure setting unit.

  Further, the pressure control valve 44 has two positions, a communication position (fully open position) for communicating the upstream side and the downstream side of the return line 42 as it is, and a throttle position for communicating in a throttled state. The opening area is continuously changed between the communication position and the throttle position by the differential pressure between the most upstream pressure Pup and the most downstream pressure Pdown of the return line 42 loaded on the portions 44a and 44b.

  FIG. 2 is an explanatory diagram of the operation of the tank return circuit 7. In the figure, ΔPb is a differential pressure between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7, and Ps is when the discharge flow rate of the main pump 2 becomes the minimum during the turning deceleration operation or the like. In addition, the pressure is such that unpleasant cavitation is not generated in the actuator oil passage 37a or 37b by quickly supplying pressure oil from the replenishment circuit 38 to the actuator oil passage 37a or 37b. Ps is, for example, about 0.3 MPa.

  The set pressure of the spring 44c of the pressure control valve 4 is set to the pressure Ps, so that the differential pressure ΔPb is independent of the flow rate Q passing through the tank return circuit 7, as shown in the lower graph of FIG. A constant pressure Ps is maintained.

  Here, the most upstream pressure Pup in the return line 42 is the back pressure of the tank return circuit 7 generated by the pressure control valve 44, and the most downstream pressure Pdown in the return line 42 is substantially equal to the tank pressure and Pdown = 0. The most upstream pressure Pup in the return line 42 (back pressure of the tank return circuit 7) is ΔPb, and this back pressure ΔPb is maintained at Ps by the pressure control valve 44.

FIG. 3 is a diagram showing the external appearance of a hydraulic excavator in which the hydraulic drive device of the present embodiment is mounted. The excavator includes a lower traveling body 101, an upper revolving body 102 that is turnably mounted on the lower traveling body 101, and a top end portion of the upper revolving body 102 that pivots vertically and horizontally via a swing post 103. And a front work machine 104 connected in a possible manner. The lower traveling body 101 is of a crawler type, and a blade 106 for earth removal is provided on the front side of the track frame 105 so as to be movable up and down. The upper swivel body 102 includes a swivel base 107 having a basic lower structure, and a canopy type cab 108 provided on the swivel base 107. The front work machine 104 includes a boom 111, an arm 112, and a bucket 113. The base end of the boom is pin-coupled to the swing post 103, and the tip of the boom 111 is pin-coupled to the base end of the arm 112. The tip is pin-coupled to the bucket 113.

  The boom 111 and the arm 112 are rotated by expanding and contracting the boom cylinder 10 and the arm cylinder 11 shown in FIG. 1, and the upper swing body 102 is rotated by rotating the swing motor 12 shown in FIG. The bucket 113 rotates by expanding and contracting the bucket cylinder 117, the blade 106 moves up and down by expanding and contracting a blade cylinder (not shown), and the lower traveling body 101 rotates the left and right traveling motors 118a and 118b. The swing post 103 rotates by expanding and contracting the swing cylinder 119. As described above, actuators such as the bucket cylinder 117, the traveling motors 118a and 118b, and the swing cylinder 119 are not shown in the hydraulic circuit diagram of FIG.

The driver's cab 108 is provided with a driver's seat 121 on which an operator is seated. A bucket / boom control lever device 122 and a swing / arm control lever device 123 are provided on the right and left sides of the driver's seat 121. A gate lock lever 124 is provided at the entrance of 121. The control lever device 122 includes a remote control valve 24 (FIG. 1), and the control lever device 123 includes remote control valves 25 and 26 (FIG. 1).
~ Operation ~
In the present embodiment, the operation will be described for each of the operation levers of the remote control valves 24 to 26 being neutral, turning steady rotation, turning deceleration, and large flow actuator operation.
<When the control lever is neutral>
The pressure oil discharged from the main pump 2 passes through the supply oil passage 2a and pressure compensation valves 35a to 35c (hereinafter represented by "35") for driving each actuator and flow control valves 34a to 34c (hereinafter "34"). To be represented). When all the operation levers of the remote control valves 24 to 26 are neutral, all the flow control valves 34 are in the neutral positions as shown in FIG.

  On the other hand, the pressure oil supplied from the main pump 2 is also led to an unload valve 31 provided between the supply oil passage 2a and the tank.

  When all the operation levers are neutral, the flow control valve 34 of each actuator is in the neutral position and the supply oil passage 2a side is closed, so that the discharge pressure (pump pressure) of the main pump 2 increases. Further, since the flow control valve 34 of each actuator is in a neutral position, the signal oil passage 27 of the maximum load pressure PLmax of each actuator 10 to 12 is connected to the tank oil passage 39 via the internal passage of each flow control valve 34. Thus, the maximum load pressure PLmax of the signal oil passage 27 becomes equal to the back pressure ΔPb (= Ps) of the tank return circuit 7. Therefore, when the discharge pressure Pd of the main pump 2 becomes higher than the sum of the set pressure of the unload valve 31 (set pressure of the spring 31a) and the back pressure ΔPb, the unload valve 31 is switched to the open position on the right side of the figure. The pressure oil in the supply oil path 2a is returned to the tank, and the discharge pressure Pd of the main pump 2 becomes equal to the sum of the set pressure of the unload valve 31 (set pressure of the spring 31a) and the back pressure ΔPb.

  Further, the differential pressure reducing valve 30 operates so as to output a differential pressure PLS between the discharge pressure Pd of the main pump 2 and the maximum load pressure PLmax of the actuators 10 to 12 as an absolute pressure. Here, the discharge pressure Pd of the main pump 2 is equal to the sum of the set pressure of the unload valve 31 (set pressure of the spring 31a) and the back pressure ΔPb, and the maximum load pressure PLmax of the actuators 10 to 12 is equal to the back pressure ΔPb. Therefore, the absolute pressure PLS is substantially equal to the set pressure of the unload valve 31 (set pressure of the spring 31a), and the unload valve 31 is set in the pressure receiving portion 4e of the LS control valve 4b of the pump tilt control device 4. An absolute pressure PLS approximately equal to the pressure is introduced. Here, as described above, the set pressure of the unload valve 31 is set to be higher than the absolute pressure Pa that is the target differential pressure of the load sensing control.

  On the other hand, the differential pressure reducing valve 5b of the engine speed detection valve device 5 outputs the differential pressure across the variable throttle portion 5c of the flow rate detection valve 5a as an absolute pressure Pa, and is received by the LS control valve 4b of the pump tilt control device 4. The absolute pressure Pa corresponding to the engine speed is guided to the part 4d as the target differential pressure.

  As described above, when the operation lever is in the neutral position, the pressure receiving portion 4d of the LS control valve 4b of the pump tilt control device 4 is guided to the absolute pressure Pa corresponding to the engine speed as the target differential pressure, and the pressure receiving portion of the LS control valve 4b. 4e is guided to an absolute pressure PLS substantially equal to the set pressure of the unload valve 31, and as a result of absolute pressure PLS> absolute pressure Pa, the LS control valve 4b of the pump tilt control device 4 is in the right position in the figure. The pressure of the pilot hydraulic power source 21 is guided to the LS control tilt actuator 4c, the tilt angle of the main pump 2 is minimized, and the discharge flow rate of the main pump 2 is also minimized.

  When the discharge flow rate of the main pump 2 is minimized, the flow rate Q returning from the control valve 6 to the tank via the tank return circuit 7 is also reduced. However, the pressure difference ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 (that is, the back pressure of the tank return circuit 7) is maintained at the optimum pressure Ps by the pressure control valve 44.

That is, when the flow rate Q is small, the pressure loss in the oil cooler 43 is small. Therefore, when the pressure control valve 44 is in the upper communication position in the drawing, the most upstream pressure Pup is not much different from the most downstream pressure Pdown. Since there is no state, the pressure control valve 44 is switched to the throttle position on the lower side of the figure by the spring force of the spring 44c, and the tank return circuit 7 is throttled. For this reason, even when the flow rate Q is small, a constant pressure Ps is ensured.
<During turning steady rotation>
Next, only the operation lever of the turning remote control valve 26 is operated to the maximum and a certain time has elapsed, that is, the upper turning body 103 (hereinafter simply referred to as “turning”) is independently rotated at a steady speed. The case will be described.

  In FIG. 1, when the operation lever of the turning remote control valve 26 is operated, the turning flow control valve 34 c of the control valve 6 is switched, and pressure oil is supplied to the turning hydraulic motor 12.

  The load pressure at the time of the steady rotation of rotation is led to the signal oil passage 27 as the maximum load pressure PLmax by the shuttle valves 33b and 33c, and this maximum load pressure PLmax is led to the differential pressure reducing valve 30 together with the discharge pressure Pd of the main pump 2. The differential pressure between the maximum load pressure PLmax and the pump discharge pressure Pd is output as the absolute pressure PLS, and the absolute pressure PLS is guided to the pressure receiving portion 4e of the LS control valve 4b of the pump tilt control device 4.

  On the other hand, the absolute pressure Pa output from the differential pressure reducing valve 5b of the engine speed detection valve device 5 is guided to the pressure receiving portion 4d of the LS control valve 4b of the pump tilt control device 4 as a target differential pressure, The tilt of the main pump 2 is controlled so that the absolute pressure PLS (the differential pressure between the maximum load pressure PLmax and the pump discharge pressure Pd) is equal to the absolute pressure Pa (target differential pressure).

  When the swing is rotating normally, the differential pressure PLmax between the maximum load pressure PLmax and the pump discharge pressure Pd is equal to the target differential pressure Pa, and the discharge pressure Pd of the main pump 2 is the load pressure PLmax of the steady swing. Than the target differential pressure Pa.

  Further, when the turn is rotating normally, the flow rate Q returning from the control valve 6 to the tank via the tank return circuit matches the flow rate discharged by the main pump 2.

  In general, in the case of turning at a steady speed, the flow rate discharged from the main pump 2, that is, the tank return flow rate Q is often slightly larger than the minimum flow rate of the main pump 2.

  On the other hand, as shown in FIG. 2, the pressure control valve 44 provided in the tank return circuit 7 has a pressure difference ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 as Ps (a constant value). ) To be kept.

That is, during steady turning, the flow rate Q is larger than when the operation lever is neutral, so the most upstream pressure Pup of the tank return circuit 7 is higher than when the operation lever is neutral. On the other hand, since the most downstream pressure Pdown does not always change equal to the tank pressure, the pressure control valve 44 has a position where the differential pressure ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown balances with the spring force of the spring 44c of the pressure control valve 44. Until the balance is reached. Therefore, the differential pressure ΔPb is maintained at Ps (a constant value).
<When turning and decelerating>
Next, an operation when the operation lever of the turning remote control valve 26 is returned by a loose operation from the state in which the upper turning body 102 is rotating in a steady manner will be described.

  In FIG. 1, when the operation lever of the remote control valve 26 for turning operation is returned to the neutral position by loose operation, the flow control valve 34c also slowly returns to the neutral position, and the pressure oil supply path to the turning hydraulic motor 12 is slowly shut off. Is done.

  Since the upper swing body 102 (FIG. 3) of a hydraulic excavator having a large moment of inertia is connected to the turning hydraulic motor 12, it rotates at the large moment of inertia until the flow control valve 34c returns to neutral. Try to continue. For this reason, high pressure is accumulated in the actuator oil passage on the side where the hydraulic oil is returned from the hydraulic motor 12 to the flow control valve 34c in the actuator oil passages 37a and 37b, and the pressure oil is supplied to the hydraulic motor 12 from the flow control valve 34c. The pressure in the actuator oil passage (swing load pressure) becomes very low.

  The swing load pressure is led from the swing flow control valve 34c to the signal oil passage 27 through the shuttle valves 33b and 33c, and this load pressure is led to the differential pressure reducing valve 30 as the maximum load pressure PLmax. Since the pressure in the actuator oil passage on the side supplying pressure oil from the flow control valve 34c to the hydraulic motor 12 becomes very low, the maximum load pressure PLmax guided to the differential pressure reducing valve 30 also becomes very low.

  On the other hand, although the flow rate flowing from the main pump 2 to the turning flow rate control valve 34c decreases, the flow rate supplied from the main pump 2 does not change instantaneously, so that the pressure oil is stored in the pressure oil supply passage 2a. The discharge pressure Pd of the pump 2 increases, and this pump discharge pressure Pd is guided to the differential pressure reducing valve 30.

  As described above, the maximum load pressure PLmax becomes very low and the pump discharge pressure Pd becomes high. As a result, the absolute pressure PLS of the differential pressure between the maximum load pressure PLmax output from the differential pressure reducing valve 30 and the pump discharge pressure Pd becomes large. The absolute pressure PLS is guided to the pressure receiving portion 4e of the LS control valve 4b of the pump tilt control device 4.

  As described above, the absolute pressure PLS guided to the pressure receiving portion 4e of the LS control valve 4b of the pump tilt control device 4 is increased, but the absolute pressure Pa () output from the differential pressure reducing valve 5b of the engine speed detection valve device 5 is increased. Since the target differential pressure is constant if the engine speed is constant, the control valve 4b is switched to the position on the right side of the figure, and the pressure of the pilot hydraulic power source 21 is guided to the LS control tilting actuator 4c. The tilt angle of the pump 2 is minimized, and the discharge flow rate of the main pump 2 is also minimized.

  When the discharge flow rate of the main pump 2 is minimized, the flow rate Q returning from the control valve 6 to the tank via the tank return circuit 7 is also reduced. However, the pressure difference ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 (that is, the back pressure of the tank return circuit 7) is maintained at the optimum pressure Ps by the pressure control valve 44.

  That is, when the flow rate Q is small, the pressure loss in the oil cooler 43 is small. Therefore, when the pressure control valve 44 is in the upper communication position in the drawing, the most upstream pressure Pup is not much different from the most downstream pressure Pdown. Since there is no state, the pressure control valve 44 is switched to the throttle position on the lower side of the figure by the spring force of the spring 44c, and the tank return circuit 7 is throttled. For this reason, even when the flow rate Q is small, a constant pressure Ps is secured.

In addition, as described above, the differential pressure ΔPb (back pressure of the tank return circuit 7) between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 is maintained at a constant pressure Ps. As a result, the flow control valve 34c. Even if the pressure of the actuator oil passage 37a or 37b on the side where pressure oil is supplied from the hydraulic motor 12 to the hydraulic motor 12 becomes very low, the replenishment of the pressure oil from the replenishment circuit 38 to the actuator oil passage 37a or 37b is performed smoothly. Unpleasant cavitation is prevented from occurring in the actuator oil passage 37a or 37b.
<When operating a large flow actuator>
Next, an operation when a large flow rate actuator such as the arm cylinder 11 that moves the arm 111 up and down is driven will be described.

  In FIG. 1, when the operation lever of the arm remote control valve 25 is operated, the arm flow control valve 34 b of the control valve 6 is switched, and pressure oil is supplied to the arm cylinder 11.

  The arm cylinder drive load pressure is led to the signal oil passage 27 as the maximum load pressure PLmax by the shuttle valves 33a and 33b, and this maximum load pressure PLmax is led to the differential pressure reducing valve 30 together with the discharge pressure Pd of the main pump 2. The differential pressure between the maximum load pressure PLmax and the pump discharge pressure Pd is output as the absolute pressure PLS, and the absolute pressure PLS is guided to the pressure receiving portion 4e of the LS control valve 4b of the pump tilt control device 4.

  On the other hand, the absolute pressure Pa output from the differential pressure reducing valve 5b of the engine speed detection valve device 5 is guided to the pressure receiving portion 4d of the LS control valve 4b of the pump tilt control device 4 as a target differential pressure, The tilt of the main pump 2 is controlled so that the absolute pressure PLS (the differential pressure between the maximum load pressure PLmax and the pump discharge pressure Pd) is equal to the absolute pressure Pa (target differential pressure).

  When the arm cylinder 11 is driven at a steady speed, the differential pressure PLmax between the maximum load pressure PLmax and the pump discharge pressure Pd is equal to the target differential pressure Pa, and the discharge pressure Pd of the main pump 2 is It is kept in a state higher than the load pressure PLmax for steady turning by the target differential pressure Pa.

  Usually, the flow rate for driving the arm cylinder 11 is often larger than the flow rate for driving the turning motor 12. Further, when the operation of contracting the arm cylinder 11 is performed, pressure oil is supplied from the rod side of the arm cylinder 11, and in this case, the arm cylinder is caused by the difference in pressure receiving area between the bottom side of the arm cylinder 11 and the rod side. The flow rate returning from 11 to the flow control valve 34b increases.

  That is, when the operation of contracting the arm cylinder 11 is performed, the flow rate returning from the arm cylinder 11 to the flow rate control valve 34b, that is, the flow rate Q returning from the control valve 6 to the tank is very large as compared with the case of the turning operation or the like. Become.

  On the other hand, as shown in FIG. 2, the pressure control valve 44 provided in the tank return circuit 7 has a pressure difference ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 (that is, the tank return). The circuit 7 operates so that the back pressure of the circuit 7 is maintained at Ps (a constant value).

  That is, when the arm cylinder 11 is contracted, the tank return flow rate Q is large, so that the pressure loss generated in the oil cooler 43 is also increased, and the most upstream pressure Pup of the tank return circuit is instantaneously increased. On the other hand, since the most downstream pressure Pdown does not always change equal to the tank pressure, the pressure control valve 44 has a pressure difference ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown overcomes the spring force of the spring 44c of the pressure control valve 44. , Switching to the upper communication position in the figure. As a result, the pressure loss generated in the pressure control valve 44 is reduced, and the pressure control valve 44 is located at a position where the differential pressure ΔPb between the most upstream pressure Pup and the most downstream pressure Pdown balances with the spring force of the spring 44c of the pressure control valve 44. To balance.

  With the above operation, even when the tank return flow rate Q is large, the differential pressure ΔPb (back pressure of the tank return circuit 7) is maintained at Ps (a constant value).

  FIG. 4 is a view showing a tank return circuit of a conventional control valve.

  The conventional tank return circuit 7X includes a back pressure valve 44X for obtaining a substantially constant back pressure regardless of the flow rate, an oil cooler 43 for cooling the hydraulic oil, and a bypass check valve connected in parallel to the oil cooler 43. 45. The back pressure valve 44X is a check valve having a certain setting spring. The back pressure valve 44X works to prevent the occurrence of cavitation because the back pressure sufficient for replenishment can be obtained even when the flow rate of the hydraulic pump is reduced by load sensing control during deceleration such as turning. it can.

  However, this conventional tank return circuit has the following problems.

  As described above, the return oil from the control valve 6 passes through the back pressure valve 44X and returns to the tank via the oil cooler 43 for cooling the hydraulic oil. The oil cooler 43 is equipped with a bypass check valve 45 in parallel with the oil cooler 43, and is opened when the oil temperature is low, the viscosity of the hydraulic oil is high, and the pressure loss of the oil cooler 43 is high. Yes. Since the pressure loss of the normal oil cooler 43 is often not so high at the oil temperature normally used, the bypass check valve 45 is closed.

The pressure loss generated by the back pressure valve 44X is ΔP ₁ , the pressure loss generated by the oil cooler 43 is ΔP ₂ , and the pressure loss generated in the entire tank return circuit is ΔPc. In this case, the relationship ΔPc = ΔP ₁ + ΔP ₂ is established. The pressure loss ΔPc corresponds to the differential pressure between the most upstream pressure Pup and the most downstream pressure Pdown of the return line 42 in FIG. 2 in the present embodiment, and is the back pressure of the tank return circuit X.

ΔP ₁ , ΔP ₂ , and ΔPc have characteristics as shown in the graphs on the lower side of FIG. That is, ΔP ₁ becomes a substantially constant pressure regardless of the flow rate Q, ΔP ₂ increases as the flow rate Q increases, and ΔPc has a characteristic obtained by adding ΔP ₁ and ΔP ₂ together.

Here, ΔP ₁ is set to a pressure Ps at which the replenishment can be performed promptly during the turning deceleration operation, that is, when the discharge flow rate of the hydraulic pump is minimized, and unpleasant cavitation does not occur.

When looking at the lower graph in FIG. 4, the pressure loss ΔPc of the entire tank return circuit acting on the control valve 6 (back pressure of the tank return circuit 7X) is almost for the prevention of cavitation when the flow rate Q is small. Although it is the same value as the required pressure Ps, when the large flow rate actuator is driven and the tank return flow rate Q is large, it becomes larger than necessary due to the effect of ΔP ₂ that increases as the flow rate Q increases. End up. This pressure loss that is larger than necessary is a wasteful energy loss, resulting in a deterioration in the energy efficiency of the hydraulic drive device.

In contrast to such a conventional technique, in the present embodiment, as described above, the pressure control valve 44 provided in the tank return circuit 7 allows the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7 to function. The pressure difference ΔPb (back pressure of the tank return circuit 7) is maintained at a constant pressure Ps regardless of the flow rate Q passing through the tank return circuit 7. As a result, even when the discharge flow rate of the main pump 2 is large, such as when a large flow rate actuator such as the arm cylinder 11 is driven, unnecessary back pressure can not be raised, and wasteful pressure loss can be reduced. The energy efficiency of the hydraulic drive device can be improved.
<Embodiment 2>
A second embodiment of the present invention will be described with reference to FIG. In the figure, the same parts as those shown in FIG.
~Constitution~
In the present embodiment, the configuration of a pressure control valve provided as a back pressure generating device in the tank return circuit 7 of the control valve 6 is different from that of the first embodiment.

  That is, in the present embodiment shown in FIG. 5, the pressure control valve 44A, which is a back pressure generator, is added to the tank return circuit 7A, as in the first embodiment, and the first pressure receiving chamber 44a acting in the opening direction. A second pressure receiving chamber 44b acting in the throttle direction located on the side facing the first pressure receiving chamber 44a, and the most upstream pressure Pup of the return line 42 is guided to the first pressure receiving chamber 44a, and the second pressure receiving chamber 44b The pressure Pdown on the most downstream side of the return line 42 is guided, and instead of the spring 44c in the first embodiment, a third pressure receiving chamber 44d having a throttle direction action located on the same side as the second pressure receiving chamber 44b; A spring 44e acting in the opening direction located on the side facing the third pressure receiving chamber 44d, and the pressure of the second pilot oil passage 22 downstream of the gate lock valve 14 is guided to the third pressure receiving chamber 44, The first generated by that pressure Pressure receiving chamber 44d of the hydraulic force and the return by the difference between the spring force of the spring 44e most upstream pressure Pup and the target differential pressure downstream of the pressure Pdown of the line 42 (the pressure Ps in FIG. 2) is set. That is, in the present embodiment, the third pressure receiving chamber 44d and the spring 44e function as target differential pressure setting means, and the pressure receiving area of the third pressure receiving chamber 44d is combined with the spring force of the opposing spring 44e, The spring 44c in the embodiment is set to have the same pressure control characteristics (so that the target differential pressure becomes the pressure Ps).

Other configurations are the same as those in the first embodiment.
~ Operation ~
This embodiment is different from the first embodiment only in the operation of the pressure control valve 44A, and the other operations are the same as those in the first embodiment. The operation of only the pressure control valve 44A will be described below.
<When the gate lock lever 124 is in an inoperable state>
When the gate lock lever 124 is in an inoperable state (raised position), as described above, the electromagnetic switching valve 14 is switched to the second position (OFF position) on the right side of the drawing, and the first pilot oil passage 20 and the second pilot are switched. The communication with the oil passage 22 is blocked, and the second pilot oil passage 22 is communicated with the tank. As a result, the pressure (tank pressure) on the downstream side of the electromagnetic switching valve 14 equal to the tank pressure is guided to the third pressure receiving chamber 44d of the pressure control valve 44A. When the tank pressure is guided to the third pressure receiving chamber 44d, the pressure control valve 44A is switched to the upper communication position in the drawing by the spring force of the spring 44e located on the side facing the third pressure receiving chamber 44d. As a result, the pressure control valve 44A does not throttle the tank return circuit 7A at all.
<When the gate lock lever is operational and all control levers are neutral>
When the gate lock lever 124 is in the operable state (lowering position), as described above, the electromagnetic switching valve 14 is switched to the first position (ON position) on the left side of the drawing, and the second pilot oil passage 22 and the tank are connected. The communication is cut off, and the first pilot oil passage 20 and the second pilot oil passage 22 are connected. As a result, the pilot pressure kept constant by the pilot relief valve 13 of the pilot hydraulic power source 21 is guided to the third pressure receiving chamber 44d of the pressure control valve 44A. When the pressure of the pilot hydraulic pressure source 21 is guided to the third pressure receiving chamber 44d, the oil pressure in the third pressure receiving chamber 44d overcomes the spring force of the opposing spring 44e, and the first embodiment is directed upward (throttle direction) in the figure. A force equivalent to that of the spring 44c is generated. As a result, the pressure difference ΔPb (back pressure of the tank return circuit 7A) between the most upstream pressure Pup and the most downstream pressure Pdown of the tank return circuit 7A is the difference between the oil pressure in the third pressure receiving chamber 44d and the spring force of the spring 44e. The pressure is controlled to be a constant pressure Ps set by the difference.
<Other cases>
In other cases, the operation of the pressure control valve 44A is the same as that of the first embodiment.

  In the present embodiment configured as described above, the following effects are obtained in addition to the effects of the first embodiment.

  As an example of making the gate lock lever 124 inoperable (up position), there is a case where the engine 1 is stopped with the engine 1 running. In such a case, in the present embodiment, when the gate lock lever 124 is disabled (in the raised position), the pressure control valve 44A is switched to the communication position. The pressure loss of the return oil circulating to the tank via the valve 31 can be suppressed to only the pressure loss generated by the oil cooler 43 in the tank return circuit 7A, and the energy efficiency can be further improved.

  Further, as another case where the gate lock lever 124 is set in an inoperable state (up position), the key switch of the engine 1 is turned off to stop the engine 1 and end one day of work. In such a case, the engine 1 is started by turning on the key switch of the engine 1 in order to resume work the next day. At this time, the engine 1 may be difficult to start if the load of the main pump 2 is large. . In the present embodiment, when the gate lock lever 124 is in an inoperable state (raised position), the pressure control valve 44A is in the communication position, and the return oil circulates from the main pump 2 to the tank via the unload valve 31. The pressure loss is only the pressure loss generated by the oil cooler 43 in the tank return circuit 7A, so the load on the main pump 2 is reduced, the load on the engine 1 is reduced, and the startability of the engine 1 is improved.

  Although one embodiment of the present invention has been described above, the present invention is not limited thereto and can be variously modified within the spirit of the present invention.

  For example, in the above-described embodiment, load sensing type pump control means is provided. However, if the capacity of the hydraulic pump is controlled according to the required flow rate of a plurality of flow control valves, other pump control is performed. It may be a means. Other pump control means include a positive control system that increases the capacity (tilt angle) of the hydraulic pump as the control pilot pressure (operation signal) of the remote control valve increases, and is located downstream of the center bypass line of the control valve. Examples include a negative control method in which a throttle is provided and the capacity (tilt angle) of the hydraulic pump is increased as the pressure on the upstream side of the throttle decreases.

  In the above embodiment, the hydraulic drive device is provided with a replenishment circuit in the actuator oil passage of the swing motor. However, the hydraulic drive device is used for other actuators (for example, a boom cylinder, an arm cylinder, a travel motor, etc.). The actuator oil passage may be provided with a replenishment circuit.

It is a figure which shows the structure of the hydraulic drive device of the construction machine (hydraulic excavator) in the 1st Embodiment of this invention. 6 is an operation explanatory diagram of a tank return circuit 7. FIG. It is a figure which shows the external appearance of the hydraulic excavator by which the hydraulic drive device of this Embodiment is mounted. It is a figure which shows the tank return circuit of the conventional control valve. It is a figure which shows the structure of the hydraulic drive device of the construction machine (hydraulic excavator) in the 2nd Embodiment of this invention.

Explanation of symbols

1 Engine 2 Main hydraulic pump (Main pump)
2a Supply oil passage 3 Pilot pump 4 Pump tilt control device 4a Horsepower control tilt actuator 4b LS control valve 4c LS control tilt actuator 4d, 4e Pressure receiving unit 5 Engine rotation speed detection valve device 5a Flow rate detection valve 5b Differential pressure reducing valve 5c Variable throttle 6 Control valve 7, 7A Tank return circuit 10 Actuator (boom cylinder)
11 Actuator (arm cylinder)
12 Actuator (Swivel motor)
13 Pilot relief valve 14 Electromagnetic switching valve 17, 18, 19 Valve pioneer 20 First pilot oil passage 21 Pilot oil pressure source 22 Second pilot oil passage 24, 25, 26 Remote control valve 27 Signal oil passage 30 Differential pressure reducing valve 31 An Load valve 31a Spring 32 Main relief valves 33a, 33b, 33c Shuttle valves 34a, 34b, 34c Flow rate control valves 35a, 35b, 35c Pressure compensation valves 36a, 36b, 36c Valve-opening pressure receiving portions 37a, 37b Actuator oil passage 38 Supply circuit 38a, 38b Check valve 39 Tank oil passage 42 Return line 43 Oil coolers 44, 44A Back pressure generator (pressure control valve)
44a First pressure receiving chamber 44b Second pressure receiving chamber 44c Spring 44d Third pressure receiving chamber 44e Spring 45 Bypass check valve 124 Gate lock lever

Claims (4)

An engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by pressure oil discharged from the hydraulic pump, and a flow rate of pressure oil supplied from the hydraulic pump to the actuators In a hydraulic drive device for a construction machine, comprising a control valve having a plurality of flow control valves for controlling the flow rate, and a pump control means for controlling the capacity of the hydraulic pump in accordance with the required flow rates of the plurality of flow control valves
A replenishment circuit provided in an actuator oil passage of at least one of the plurality of actuators;
A tank having a return line for circulating return oil from the control valve to the tank, an oil cooler and a back pressure generator provided in the return line, and the replenishment circuit connected to the upstream side of the back pressure generator A return circuit,
The hydraulic drive device for a construction machine, wherein the back pressure generating device is a pressure control valve that controls the pressure difference between the most upstream pressure and the most downstream pressure in the return line to be constant. An engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by pressure oil discharged from the hydraulic pump, and a flow rate of pressure oil supplied from the hydraulic pump to the actuators A control valve having a plurality of flow control valves for controlling the flow rate, and a load sensing type pump for controlling the capacity of the hydraulic pump so that a discharge pressure of the hydraulic pump is higher than a maximum load pressure of the plurality of actuators by a target differential pressure In a hydraulic drive device for a construction machine comprising a control means,
A replenishment circuit provided in an actuator oil passage of at least one of the plurality of actuators;
A tank having a return line for circulating return oil from the control valve to the tank, an oil cooler and a back pressure generator provided in the return line, and the replenishment circuit connected to the upstream side of the back pressure generator A return circuit,
The hydraulic drive device for a construction machine, wherein the back pressure generating device is a pressure control valve that controls the pressure difference between the most upstream pressure and the most downstream pressure in the return line to be constant. The hydraulic drive device for a construction machine according to claim 1 or 2,
A pilot hydraulic source,
Provided on the downstream side of the pilot hydraulic source, the pressure of the pilot hydraulic source is output downstream when the gate lock lever is in an operable state, and the tank pressure is output when the gate lock lever is in an inoperable state. A gate lock valve that outputs to the downstream side;
The pressure control valve includes a first pressure receiving chamber acting in an opening direction, second and third pressure receiving chambers acting in a throttle direction, and spring means acting in an opening direction, and the first pressure receiving chamber has an outermost return line. An upstream pressure is guided, a pressure downstream of the return line is guided to the second pressure receiving chamber, a pressure downstream of the gate lock valve is guided to the third pressure receiving chamber, and a pressure generated by the pressure is generated. 3. A hydraulic drive device for a construction machine, wherein a target differential pressure between the most upstream pressure and the most downstream pressure in the return line is set by the difference between the oil pressure in the pressure receiving chamber and the spring force of the spring means. The hydraulic drive device for a construction machine according to claim 1 or 2,
The pressure control valve has a first pressure receiving chamber acting in the opening direction, a second pressure receiving chamber acting in the throttle direction, and a spring means acting in the throttle direction, and the pressure of the most upstream of the return line in the first pressure receiving chamber. The pressure downstream of the return line is guided to the second pressure receiving chamber, and the target differential pressure between the most upstream pressure and the most downstream pressure of the return line is set by the spring means. Hydraulic drive device for construction machinery.

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