JP2016169818A - Hydraulic driving system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hydraulic driving system capable of bringing an operation point close to an engine maximum output torque line in a high region from an intermediate region of an engine rotating speed without using an electromagnetic proportional valve.SOLUTION: A hydraulic driving system 1A includes a variable capacity type main pump 15 driven by an engine for supplying a hydraulic oil to a hydraulic actuator, a fixed capacity type sub pump 19 driven by the engine, a throttle 62 disposed on a sub pump discharge line 61 extending from the sub pump, a relief valve 63 disposed on the sub pump discharge line at a downstream side of the throttle, and a regulator which is a regulator 16 changing an inclination angle of the main pump, and including a horse power control piston 83 receiving a discharge pressure of the main pump, and a reaction piston 87 operated against the horse power control piston by receiving a pressure at an upstream side of the throttle in the sub pump discharge line.SELECTED DRAWING: Figure 3

Description

  The present invention relates to a hydraulic drive system mounted on, for example, a construction machine.

  Some industrial machines and construction machines are equipped with a hydraulic drive system including a variable displacement pump driven by an engine. For example, Patent Document 1 discloses a hydraulic drive system for a hydraulic excavator.

  The hydraulic drive system disclosed in Patent Document 1 drives a variable capacity first main pump and a second main pump, and a first main pump and a second main pump for supplying hydraulic oil to a plurality of hydraulic actuators. Including the engine to do. The tilt angle of the first main pump is changed by the first regulator, and the tilt angle of the second main pump is changed by the second regulator.

  Each of the first regulator and the second regulator includes a first servo valve for positive tilt control and a second servo valve for full horsepower control. The first servo valve is operated according to the secondary pressure output from the first electromagnetic proportional valve, and the second servo valve is the discharge pressure of the first main pump, the discharge pressure of the second main pump, and the second electromagnetic proportional. Operates according to the secondary pressure output from the valve. The first electromagnetic proportional valve and the second electromagnetic proportional valve are connected to a sub-pump driven by the engine.

  The second electromagnetic proportional valve outputs a secondary pressure to both the second servo valve of the first regulator and the second servo valve of the second regulator. The second electromagnetic proportional valve is for adjusting the absorption (load) torque of the first and second main pumps according to the engine speed.

  Generally, the hydraulic drive system is provided with a rotation speed selection device for the operator to select a reference engine rotation speed in several stages. FIG. 7 shows a linear engine droop line EL2 determined for each reference engine speed. The engine torque during operation changes on a curve defined by the engine droop line EL2 and the engine maximum output torque line EL1 corresponding to the selected reference engine speed.

  The second electromagnetic proportional valve described above determines a pump absorption (load) torque line PL by horsepower control shown in FIG. That is, the second electromagnetic proportional valve is controlled by the control device so that the absorption torque of the first and second main pumps also decreases when the engine speed decreases. The intersection of the pump absorption torque line PL and the engine maximum output torque line EL1 or the engine droop line EL2 is an actual operating point.

JP-A-11-101183

  In the hydraulic drive system disclosed in Patent Document 1, the absorption (load) torque of the first and second main pumps is adjusted using an electromagnetic proportional valve. However, when an electromagnetic proportional valve is used, a failure due to poor contact of electrical components (connectors, harnesses, etc.) is likely to occur, particularly in regions where the usage environment is severe.

  From such a viewpoint, it is conceivable to delete the second electromagnetic proportional valve from the hydraulic drive system disclosed in Patent Document 1. In this case, as shown in FIG. 8, the pump absorption torque by the horsepower control is constant (in other words, the pump absorption torque line PL is horizontal). For this reason, it is necessary to keep the pump absorption torque low in order to prevent the operating point from reaching the maximum torque point on the engine maximum output torque line EL1 no matter which reference engine speed is selected.

  However, when the ends of the engine maximum output torque line EL1 are compared with each other, the end of the region higher than the intermediate region of the engine speed has a larger torque than the end of the lower engine speed, so In a region higher than the region, there is a margin from the operating point to the engine maximum output torque line EL1.

  Therefore, an object of the present invention is to provide a hydraulic drive system that can bring the operating point close to the engine maximum output torque line in a region higher than the intermediate region of the engine speed without using an electromagnetic proportional valve.

  In order to solve the above problems, a hydraulic drive system according to the present invention includes a variable displacement main pump driven by an engine for supplying hydraulic oil to a hydraulic actuator, and a fixed displacement drive driven by the engine. A sub-pump, a throttle provided in a sub-pump discharge line extending from the sub-pump, a relief valve provided in the sub-pump discharge line downstream of the throttle, and a regulator that changes a tilt angle of the main pump, A regulator including a horsepower control piston that receives the discharge pressure of the main pump, and a reaction force piston that operates against the horsepower control piston by receiving pressure on the upstream side of the throttle in the sub-pump discharge line. It is characterized by that.

  According to the above configuration, when the engine speed increases, the discharge flow rate of the sub pump also increases, and the pressure upstream of the throttle in the sub pump discharge line increases. Since the pressure that rises as the engine speed increases in this way acts on the reaction force piston that operates against the horsepower control piston, the restriction of the pump horsepower by the horsepower control piston is reduced as the engine speed increases. Is done. As a result, a pump absorption torque line that is proportional to the engine speed can be obtained. As a result, the operating point can be brought close to the engine maximum output torque line in a region higher than the intermediate region of the engine speed.

  The hydraulic drive system includes a control valve that is disposed on a circulation line that extends from the main pump to the tank and that controls supply and discharge of hydraulic oil to and from the hydraulic actuator, and a pilot operation valve that operates the control valve. And an operation valve supply line branched from the sub-pump discharge line downstream of the throttle and connected to the pilot operation valve. According to this configuration, since the sub pump functions as a pilot pump, the pilot pump can be used to generate a pressure that increases as the engine speed increases. Therefore, there is no need to provide a pump dedicated to the reaction force piston.

  The circulation line may be provided with a throttle on the downstream side of the control valve, and the regulator may include a flow control piston that receives a negative control pressure that is a pressure upstream of the throttle in the circulation line. According to this configuration, since the discharge flow rate of the main pump can be controlled by the hydraulic negative control method, there are fewer electrical components than the electrical positive control method that requires an electromagnetic proportional valve for flow rate control. Therefore, the cost can be reduced and the hydraulic drive system can be stably operated in a harsh environment.

  The regulator includes a first horsepower control chamber and a second horsepower control chamber facing the horsepower control piston, and the discharge pressure of the main pump is guided to the first horsepower control chamber. A relay line connecting the second horsepower control chamber and a portion of the sub-pump discharge line between the throttle and the relief valve; and a pilot line provided in the relay line when the pilot pressure is a predetermined value or less. A two-horsepower control chamber in communication with the tank, and when the pilot pressure is greater than the predetermined value, a pressure reducing valve that outputs a secondary pressure proportional to the pilot pressure to the second horsepower control chamber; and the sub-pump discharge line A pressure reducing valve signal input line that guides the pressure upstream of the throttle to the pressure reducing valve as the pilot pressure. Good. According to this configuration, the pump absorption torque line can be a broken line that peaks at the engine speed corresponding to the predetermined value. As a result, the operating point can be made closer to the engine maximum output torque line in the intermediate region of the engine speed.

  According to the present invention, the operating point can be brought close to the engine maximum output torque line in a region higher than the intermediate region of the engine speed without using an electromagnetic proportional valve.

1 is an overall schematic configuration diagram of a hydraulic drive system according to a first embodiment of the present invention. It is a side view of the hydraulic excavator which is an example of a construction machine. It is a partial schematic block diagram of the hydraulic drive system which concerns on 1st Embodiment. (A) is a graph which shows the relationship between the engine speed in 1st Embodiment, and an engine torque, (b) is a graph which shows the relationship between the engine speed in 1st Embodiment, and the discharge pressure of a subpump. It is a partial schematic block diagram of the hydraulic drive system which concerns on 2nd Embodiment of this invention. (A) is a graph which shows the relationship between the engine speed and engine torque in 1st Embodiment, (b) is the relationship between the engine speed in 1st Embodiment, the discharge pressure of a subpump, and the output pressure of a pressure-reduction valve. It is a graph which shows. It is a graph which shows the relationship between the engine speed and engine torque in the conventional hydraulic drive system. It is a graph which shows the relationship between the engine speed and engine torque in the hydraulic drive system of a virtual example.

(First embodiment)
FIG. 1 shows a hydraulic drive system 1A for a construction machine according to a first embodiment of the present invention, and FIG. 2 shows a construction machine 10 on which the hydraulic drive system 1A is mounted. The construction machine 10 shown in FIG. 2 is a hydraulic excavator, but the present invention is also applicable to other construction machines such as a hydraulic crane. The present invention can be applied not only to construction machines but also to other machines (for example, industrial machines).

  The hydraulic drive system 1A includes a boom cylinder 11, an arm cylinder 12, and a bucket cylinder 13 shown in FIG. 2 as hydraulic actuators, and also includes a turning motor 14 shown in FIG. 1 and a pair of left and right traveling motors (not shown). The hydraulic drive system 1A includes a first main pump 15 and a second main pump 17 for supplying hydraulic oil to those actuators, and an engine 51 for driving the first main pump 15 and the second main pump 17. . In FIG. 1, actuators other than the boom cylinder 11 and the swing motor 14 are omitted for simplification of the drawing.

  The engine 51 includes a fuel injection valve that is controlled by a governor 52. The governor 52 is connected to a rotation speed selection device 53 for the operator to select the reference engine rotation speed in several stages. In other words, the rotation speed selection device 53 receives selection of the reference engine rotation speed. The governor 52 controls the fuel injection valve according to the selected reference engine speed.

  A first circulation line 21 extends from the first main pump 15 to the tank. On the first circulation line 21, a plurality of control valves including the turning control valve 41 (not shown except for the turning control valve 41) are arranged. Control valves other than the turning control valve 41 are, for example, an arm control valve and a travel left control valve. The swing control valve 41 controls the supply and discharge of hydraulic oil to the swing motor 14, and the other control valves also control the supply and discharge of hydraulic oil to the individual actuators. A parallel line 24 branches off from the first circulation line 21, and hydraulic oil discharged from the first main pump 15 is guided to all control valves on the first circulation line 21 through the parallel line 24. A tank line 25 is connected to each control valve on the first circulation line 21.

  Similarly, the second circulation line 31 extends from the second main pump 17 to the tank. On the second circulation line 31, a plurality of control valves including the boom control valve 43 (not shown except for the boom control valve 43) are arranged. Control valves other than the boom control valve 43 are, for example, a bucket control valve and a traveling right control valve. The boom control valve 43 controls the supply and discharge of hydraulic fluid to the boom cylinder 11, and the other control valves also control the supply and discharge of hydraulic fluid to individual actuators. A parallel line 34 branches from the second circulation line 31, and hydraulic oil discharged from the second main pump 17 is guided to all control valves on the second circulation line 31 through the parallel line 34. A tank line 35 is connected to each control valve on the second circulation line 31.

  Further, the hydraulic drive system 1 </ b> A includes a swing operation valve 42 that is a pilot operation valve for operating the swing control valve 41, and a boom operation valve 44 that is a pilot operation valve for operating the boom control valve 43. Each of the turning operation valve 42 and the boom operation valve 44 has an operation lever, and outputs a pilot pressure having a magnitude corresponding to the tilt angle (operation amount) of the operation lever to the turning control valve 41 or the boom control valve 43.

  The primary pressure is supplied to the swing operation valve 42 and the boom operation valve 44 from the fixed displacement sub pump 19 driven by the engine 51. As shown in FIG. 3, a sub pump discharge line 61 extends from the sub pump 19 to the tank. The sub pump discharge line 61 is provided with a throttle 62 and a relief valve 63 on the downstream side of the throttle 62. An operation valve supply line 64 branches off from the portion between the throttle 62 and the relief valve 63 in the sub pump discharge line 61, and the operation valve supply line 64 is connected to the turning operation valve 42 and the boom operation valve 44.

  Returning to FIG. 1, each of the first main pump 15 and the second main pump 17 is a variable displacement pump (swash plate pump or oblique shaft pump) whose tilt angle can be changed. The tilt angle of the first main pump 15 is changed by the first regulator 16, and the tilt angle of the second main pump 17 is changed by the second regulator 18. In the present embodiment, the discharge flow rates of the first main pump 15 and the second main pump 17 are controlled by a hydraulic negative control method. However, the discharge flow rates of the first and second main pumps 15 and 17 may be controlled by a load sensing method.

  Specifically, the first circulation line 21 is provided with a throttle 22 on the downstream side of all the control valves. Further, a bypass line that bypasses the throttle 22 is connected to the first circulation line 21, and a relief valve 23 is disposed on the bypass line. Similarly, the second circulation line 31 is provided with a throttle 32 on the downstream side of all the control valves. Further, a bypass line that bypasses the throttle 32 is connected to the second circulation line 31, and a relief valve 33 is disposed on the bypass line.

  A first negative control pressure, which is a pressure upstream of the throttle 22 in the first circulation line 21, is guided to the first regulator 16 described above through the first flow rate control line 27. Further, the discharge pressure of the first main pump 15 is guided to the first regulator 16 through the first horsepower control line 26. In the present embodiment, cross sensing is not employed, and the discharge pressure of the second main pump 17 is not guided to the first regulator 16. However, cross sensing may be employed, and the discharge pressure of the second main pump 17 may be guided to the first regulator 16.

  Similarly, a second negative control pressure that is a pressure upstream of the throttle 32 in the second circulation line 31 is led to the second regulator 18 through the second flow rate control line 37. Further, the discharge pressure of the second main pump 17 is guided to the second regulator 18 through the second horsepower control line 36. In the present embodiment, cross sensing is not employed, and the discharge pressure of the first main pump 15 is not guided to the second regulator 18. However, cross sensing may be employed, and the discharge pressure of the first main pump 15 may be guided to the second regulator 18.

  In the present specification, the term “horsepower control” is used in the sense of torque control of the pump, as usual.

  As the flow rate control, the first regulator 16 reduces the tilt angle of the first main pump 15 if the first negative control pressure is high, and increases the tilt angle of the first main pump 15 if the first negative control pressure is low. To do. Further, the first regulator 16 reduces the tilt angle of the first main pump 15 if the discharge pressure of the first main pump 15 is high and controls the first main pump if the discharge pressure of the first main pump 15 is low. The tilt angle of the pump 15 is increased. When the tilt angle of the first main pump 15 decreases, the discharge flow rate of the first main pump 15 decreases, and when the tilt angle of the first main pump 15 increases, the discharge flow rate of the first main pump 15 increases.

  Similarly, as the flow rate control, the second regulator 18 reduces the tilt angle of the second main pump 17 if the second negative control pressure is high, and tilts the second main pump 17 if the second negative control pressure is low. Increase the corner. In addition, the second regulator 18 reduces the tilt angle of the second main pump 17 if the discharge pressure of the second main pump 17 is high, and reduces the second main pump 17 if the discharge pressure of the second main pump 17 is low. The tilt angle of the pump 17 is increased. When the tilt angle of the second main pump 17 decreases, the discharge flow rate of the second main pump 17 decreases, and when the tilt angle of the second main pump 17 increases, the discharge flow rate of the second main pump 17 increases.

  The first regulator 16 and the second regulator 18 have the same configuration as shown in FIG. Therefore, in the following, the configuration of the first regulator 16 will be described as a representative.

  The first regulator 16 includes a servo cylinder 72 that determines the tilt angle of the first main pump 15 and a switching valve 76 that operates the servo cylinder 72. For example, when the first main pump 15 is a swash plate pump, the servo cylinder 72 slides in the axial direction in conjunction with the swash plate 71 of the first main pump 15 and the angle of the swash plate 71. Are linked together. The first regulator 16 includes a first servo chamber 73 that faces the end of the servo cylinder 72 on the small diameter side, and a second servo chamber 74 that faces the end of the servo cylinder 72 on the large diameter side. The discharge pressure of the first main pump 15 is guided to the first servo chamber 73, and the second servo chamber 74 is connected to the switching valve 76.

  The switching valve 76 controls the opening area of the second servo chamber 74 with respect to the first circulation line 21 and the opening area with respect to the tank. Specifically, the switching valve 76 includes a sleeve 78 connected to the servo cylinder 72 by the first lever 75, and a spool 77 accommodated in the sleeve 78. For example, the axial direction of the sleeve 78 and the spool 77 is parallel to the axial direction of the servo cylinder 72. The sleeve 78 slides in the axial direction in conjunction with the attitude of the first lever 75, and the sleeve 78 relative to the spool 77 is balanced so that the force (pressure × servo cylinder pressure receiving area) acting from both sides of the servo cylinder 72 is balanced. The position is adjusted.

  The first regulator 16 includes a flow control piston 81 and a horsepower control piston 84 that drive the spool 77 of the switching valve 76. The flow control piston 81 and the horsepower control piston 84 slide in the axial direction of the spool 77 in conjunction with the posture of the second lever 83 or the third lever 86 via the second lever 83 and the third lever 86, respectively. Are connected to the spool 77. Further, the first regulator 16 includes a flow control chamber 82 facing the flow control piston 81 and a horsepower control chamber 85 facing the horsepower control piston 84.

  A first negative control pressure is introduced into the flow control chamber 82. The flow rate control piston 81 receives the first negative control pressure, and moves the spool 77 in the flow rate decreasing direction (the direction in which the discharge flow rate of the first main pump 15 decreases) when the first negative control pressure rises. When the negative control pressure decreases, the spool 77 is moved in the flow rate increasing direction (the direction in which the discharge flow rate of the first main pump 15 increases). The discharge pressure of the first main pump 15 is guided to the horsepower control chamber 85. The horsepower control piston 84 receives the discharge pressure of the first main pump 15, moves the spool 77 in the flow rate reduction direction when the discharge pressure of the first main pump 15 increases, and the discharge pressure of the first main pump 15 decreases. When this occurs, the spool 77 is moved in the direction of increasing the flow rate. Note that the flow rate control piston 81 and the horsepower control piston 84 are configured such that the one that restricts (reduces) the discharge flow rate of the first main pump 15 functions preferentially.

  Furthermore, in this embodiment, the structure for adjusting the absorption torque of the 1st main pump 15 and the 2nd main pump 17 is employ | adopted without using an electromagnetic proportional valve. Specifically, each of the first regulator 16 and the second regulator 18 includes a reaction force piston 87 that operates against the horsepower control piston 84 and a reaction force chamber 88 that faces the reaction force piston 87.

  The reaction force chamber 88 is connected to the upstream portion of the throttle 62 in the sub pump discharge line 61 by a reaction force line 65. For this reason, the reaction force piston 87 receives the counter pressure Pz that is the pressure on the upstream side of the throttle 62 in the sub-pump discharge line 61, and pushes back the horsepower control piston 84 in accordance with the counter pressure Pz.

  Since the sub pump discharge line 61 is provided with a throttle 62, if the engine speed increases, the discharge flow rate of the sub pump 19 also increases and the counter pressure Pz increases as shown in FIG. 4B. Thus, the counter pressure Pz that rises with the increase in the engine speed acts on the reaction force piston 87 that operates against the horsepower control piston 84, and therefore, the limitation of the pump horsepower by the horsepower control piston 84 is the engine speed. As the value increases, it is relaxed. Thereby, as shown to Fig.4 (a), the pump absorption torque line PL which becomes a proportional relationship with an engine speed can be obtained. As a result, the operating point (intersection of the engine droop line EL2 corresponding to the selected reference engine speed and the pump absorption torque line PL) is brought closer to the engine maximum output torque line EL1 in a region higher than the intermediate region of the engine speed. Can do.

  In the present embodiment, the discharge flow rates of the first main pump 15 and the second main pump 17 can be controlled by the hydraulic negative control method, and therefore, compared with the electrical positive control method that requires an electromagnetic proportional valve for the flow rate control. Less is. Therefore, the cost can be reduced and the hydraulic drive system 1A can be stably operated in a harsh environment. However, the present invention is not necessarily limited to the hydraulic negative control system, and the secondary pressure may be output from the electromagnetic proportional valve to the flow rate control chamber 82.

  In the embodiment, the primary pressure is supplied from the sub pump 19 to the operation valves 42 and 44. However, a pilot pump dedicated to the operation valves 42 and 44 and a sub pump dedicated to the reaction force piston 87 may be provided separately. However, with the configuration as in the above embodiment, the sub pump 19 functions as a pilot pump. Therefore, the pilot pump can be used to generate a pressure that increases as the engine speed increases. Therefore, there is no need to provide a pump dedicated to the reaction force piston.

(Second Embodiment)
Next, a hydraulic drive system 1B according to a second embodiment of the present invention will be described with reference to FIG. 5 and FIGS. 6 (a) and 6 (b). In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and a duplicate description is omitted.

  The overall configuration of the hydraulic drive system 1B is the same as that shown in FIG. In the present embodiment, each of the first regulator 16 and the second regulator 18 includes the horsepower control chamber 85 described in the first embodiment as the first horsepower control chamber 85 and the second horsepower facing the horsepower control piston 84. A control chamber 91 is included.

  The second horsepower control chamber 91 is connected to a portion between the throttle 62 and the relief valve 63 in the sub pump discharge line 61 by a relay line 92. Note that one end of the relay line 92 is not necessarily connected directly to the sub-pump discharge line 61, and may be connected to the sub-pump discharge line 61 via the operation valve supply line 64.

  The relay line 92 is provided with a pressure reducing valve 93, and a throttle 94 is provided between the pressure reducing valve 93 and the sub-pump discharge line 61 for securing a primary pressure for the operation valves 42 and 44. A tank line 95 is connected to the pressure reducing valve 93, and the pressure reducing valve 93 is configured to operate in accordance with the pilot pressure. Specifically, the pressure reducing valve 93 does not operate when the pilot pressure is equal to or less than the predetermined value α, and allows the second horsepower control chamber 91 to communicate with the tank. On the other hand, when the pilot pressure is larger than the predetermined value α, the pressure reducing valve 93 outputs a secondary pressure Pf proportional to the pilot pressure to the second horsepower control chamber 91 as shown in FIG.

  The pilot port of the pressure reducing valve 93 is connected to the upstream portion of the throttle 62 in the sub pump discharge line 61 by a pressure reducing valve signal input line 96. That is, the counter pressure Pz that is the pressure on the upstream side of the throttle 62 in the sub-pump discharge line 61 is introduced to the pilot port of the pressure reducing valve 93 as the pilot pressure.

  In the present embodiment, as the pressure acting on the horsepower control piston 84, the secondary pressure Pf output from the pressure reducing valve 93 is superimposed on the discharge pressure Pd of the main pump (the first main pump 15 or the second main pump 17). be able to. As a result, as shown in FIG. 6A, the pump absorption torque line PL can be a polygonal line that peaks at the engine speed corresponding to the predetermined value α. That is, on the left side of the broken line in FIG. 6A, the slope of the pump absorption torque line PL is proportional to the coefficient Pz−C1 × Pd (C1 is a constant), and on the right side of the broken line, the slope of the pump absorption torque line PL. Is proportional to the coefficient Pz− (C1 × Pd + C2 × Pf) (C2 is a constant). As a result, the operating point can be made closer to the engine maximum output torque line EL1 in the intermediate region of the engine speed.

(Other embodiments)
The present invention is not limited to the first and second embodiments described above, and various modifications can be made without departing from the scope of the present invention.

1A, 1B Hydraulic drive system 11 Boom cylinder (hydraulic actuator)
14 Swing motor (hydraulic actuator)
15, 17 Main pump 16, 18 Regulator 19 Sub pump 21, 31 Circulation line 22, 32 Restriction 41, 43 Control valve 42, 44 Pilot operation valve 61 Sub pump discharge line 62 Restriction 63 Relief valve 64 Operation valve supply line 81 Flow control piston 83 Horsepower control piston 85 Horsepower control chamber, first horsepower control chamber 87 Reaction force piston 91 Second horsepower control chamber 92 Relay line 93 Pressure reducing valve 96 Pressure reducing valve signal input line

Claims (4)

A variable displacement main pump driven by an engine for supplying hydraulic oil to a hydraulic actuator;
A fixed displacement sub-pump driven by the engine;
A throttle provided in a sub-pump discharge line extending from the sub-pump;
A relief valve provided in the sub-pump discharge line downstream of the throttle;
A regulator for changing the tilt angle of the main pump, which resists the horsepower control piston by receiving a pressure on the upstream side of the throttle in the sub pump discharge line and a horsepower control piston receiving the discharge pressure of the main pump. A regulator including a reaction force piston,
A hydraulic drive system comprising: A control valve disposed on a circulation line extending from the main pump to the tank and controlling supply and discharge of hydraulic oil to and from the hydraulic actuator;
A pilot operated valve for operating the control valve;
2. The hydraulic drive system according to claim 1, further comprising an operation valve supply line branched from the sub-pump discharge line downstream of the throttle and connected to the pilot operation valve. The circulation line is provided with a throttle on the downstream side of the control valve,
The hydraulic drive system according to claim 2, wherein the regulator includes a flow control piston that receives a negative control pressure that is a pressure upstream of the throttle in the circulation line. The regulator includes a first horsepower control chamber and a second horsepower control chamber facing the horsepower control piston, and a discharge pressure of the main pump is guided to the first horsepower control chamber,
A relay line connecting the second horsepower control chamber and a portion of the sub-pump discharge line between the throttle and the relief valve;
When the pilot pressure is less than or equal to a predetermined value, the second horsepower control chamber provided in the relay line communicates with the tank, and when the pilot pressure is greater than the predetermined value, a secondary pressure proportional to the pilot pressure is set A pressure reducing valve that outputs to the second horsepower control chamber;
The hydraulic drive system according to any one of claims 1 to 3, further comprising: a pressure reducing valve signal input line that guides the pressure upstream of the throttle in the sub-pump discharge line to the pressure reducing valve as the pilot pressure.

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