JP6525898B2 - Hydraulic drive of construction machine - Google Patents

Description

  The present invention relates to a hydraulic drive system for a construction machine such as a hydraulic shovel, and more particularly to a pump control apparatus (regulator) including at least two variable displacement hydraulic pumps, one of which hydraulic pumps performs at least torque control. The present invention relates to a hydraulic drive system for a construction machine having a pump controller (regulator) having load sensing control and torque control.

In a hydraulic drive system of a construction machine such as a hydraulic shovel, the flow rate of the hydraulic pump is controlled by controlling the displacement of the hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of the plurality of actuators by the target differential pressure. Those equipped with a regulator are widely used, and this control is called load sensing control. In Patent Document 1, in a hydraulic drive system of a construction machine provided with a regulator that performs such load sensing control, there are provided two hydraulic pumps and two pumps in which load sensing control is performed by each of the two hydraulic pumps. A load sensing system is described.

  Also, in the regulator of the hydraulic drive system of the construction machine, the hydraulic pump's absorption torque usually does not exceed the rated output torque of the prime mover by decreasing the hydraulic pump's capacity as the discharge pressure of the hydraulic pump increases. Control is performed to prevent the engine from stopping due to overtorque (engine stall). When the hydraulic drive includes two hydraulic pumps, the regulator of one hydraulic pump performs torque control by taking in parameters relating to the absorption torque of the other hydraulic pump as well as its own discharge pressure (total torque control), We aim at prevention of stop of motor and effective use of rated output torque of motor.

  For example, in Patent Document 2, the full torque control is performed by guiding the discharge pressure of one hydraulic pump to the regulator of the other hydraulic pump via a pressure reducing valve. The set pressure of the pressure reducing valve is constant, and the set pressure is set to a value simulating the maximum torque of torque control of the regulator of the other hydraulic pump. Thus, in the operation of driving only the actuator related to one hydraulic pump, one hydraulic pump can effectively use substantially all of the rated output torque of the prime mover, and simultaneously drives the actuator related to the other hydraulic pump In combined operation work, the absorption torque of the entire pump does not exceed the rated output torque of the prime mover, and the stop of the prime mover can be prevented.

  In Patent Document 3, in order to perform full torque control on two variable displacement hydraulic pumps, two main and sub variable pressure reducing valves are used, and the output pressure of the main variable pressure reducing valve is By making the load sensing drive pressure variable, the control accuracy of all torque control is improved. Also, the load sensing drive pressure of the other hydraulic pump is introduced to the main variable pressure reducing valve through the sub variable pressure reducing valve, and the output pressure of the sub variable pressure reducing valve is made variable by the discharge pressure of the other hydraulic pump. The change in the absorption torque of the other hydraulic pump when the other hydraulic pump is at the minimum displacement is simulated to perform all torque control with high accuracy.

  In patent document 4, in order to perform full torque control with respect to two variable displacement hydraulic pumps, the variable relief valve which changes the output pressure of an auxiliary pump according to the change of tilting angle of the other hydraulic pump is provided. By changing the discharge pressure of the auxiliary pump by this variable relief valve, the tilt angle of the other hydraulic pump is detected as the discharge pressure of the auxiliary pump, and the discharge pressure is led to the regulator of one hydraulic pump.

JP, 2011-196438, A Patent No. 3865590 JP, 2015-148236, A Japanese Examined Patent Publication 3-7030

  By incorporating the technology of total torque control described in Patent Document 2 into the two pump load sensing system described in Patent Document 1, it is possible to perform all torque control also in the two pump load sensing system described in Patent Document 1 become. However, in the total torque control of Patent Document 2, as described above, the set pressure of the pressure reducing valve is set to a constant value simulating the maximum torque of the torque control of the other hydraulic pump. For this reason, when the other hydraulic pump is subjected to the torque control limitation and is in an operating state where it operates at the maximum torque control torque in the combined operation that simultaneously drives the actuators related to the two hydraulic pumps, the motor rating Effective use of output torque can be achieved. However, when the other hydraulic pump is not restricted by torque control and is in an operating state where capacity control is performed by load sensing control, the absorption torque of the other hydraulic pump is smaller than the maximum torque of torque control. The output pressure of the pressure reducing valve simulating the maximum torque is led to the regulator of one hydraulic pump, and control is performed to reduce the absorption torque of one hydraulic pump more than necessary. For this reason, all torque control could not be performed accurately.

  In Patent Document 3, in order to perform full torque control on two variable displacement hydraulic pumps, two main and sub variable pressure reducing valves are used, and the output pressure of the main variable pressure reducing valve is The load sensing drive pressure of the other hydraulic pump is made variable by the load sensing drive pressure, and the load sensing drive pressure of the other hydraulic pump is led to the main variable pressure reducing valve through the sub variable pressure reducing valve, and the output pressure of the sub variable pressure reducing valve is By making it variable by the discharge pressure, the control accuracy of all torque control is improved. However, since the load sensing drive pressures of the two variable pressure reducing valves and the other hydraulic pump are used, there is a problem that the circuit configuration becomes complicated and the apparatus also becomes large. In particular, in the case of a small hydraulic shovel and a so-called small rear swing type having a small rear end radius, the space for storing the hydraulic pump may be small and mounting may be difficult.

  In Patent Document 4, the tilt angle of the other hydraulic pump is detected as the discharge pressure of the auxiliary pump by a variable relief valve, and the discharge pressure is introduced to the regulator of one hydraulic pump to improve the accuracy of the total torque control. And However, in general, the torque of the pump is determined by the product of the discharge pressure and the displacement, that is, (discharge pressure × pump volume) / 2π, whereas in Patent Document 4, the discharge pressure of one hydraulic pump is Lead to one of the two pilot chambers, guide the output pressure of the pressure reducing valve (proportional discharge pressure of the other hydraulic pump) to the other pilot chamber of the stepped piston, and sum the discharge pressure and discharge proportional pressure to the output torque Since the displacement of one hydraulic pump is controlled as the parameter of (1), there has been a problem that a considerable error occurs with the torque actually used.

  An object of the present invention is a construction machine having a pump control device in which one hydraulic pump performs at least torque control, and another hydraulic pump having at least two variable displacement hydraulic pumps performing load sensing control and torque control. In the hydraulic drive system, the absorption torque of the other hydraulic pump is accurately detected with a simple configuration and fed back to one hydraulic pump side, thereby performing all torque control with high accuracy and effectively utilizing the rated output torque of the prime mover To provide a hydraulic drive that can

In order to achieve the above object, the present invention provides a motor, a variable displacement first pump device driven by the motor, a variable displacement second pump device driven by the motor, the first and the second A plurality of actuators driven by pressure oil discharged by a second pump device, a plurality of flow control valves controlling flow rates of pressure oil supplied from the first and second pump devices to the plurality of actuators, A plurality of pressure compensation valves for controlling the differential pressure of the plurality of flow control valves respectively, a first pump control device for controlling the discharge flow rate by controlling the capacity of the first pump device, and the second pump device comprising of a second pump control device for controlling the discharge flow rate by controlling the capacitance, the first pump controller, at least one of discharge pressure and capacity of the first pump unit is increased, A first torque control unit configured to control a capacity of the first pump device such that the absorption torque of the first pump device does not exceed a first maximum torque when the absorption torque of the first pump device increases; In the second pump control device, when at least one of the discharge pressure and the volume of the second pump device increases and the absorption torque of the second pump device increases, the absorption torque of the second pump device is the second maximum. A second torque control unit for controlling the displacement of the second pump device so as not to exceed the torque, and a discharge pressure of the second pump device when the absorption torque of the second pump device is smaller than the second maximum torque Load sensing for controlling the capacity of the second pump device such that the target differential pressure is higher than the maximum load pressure of the actuator driven by the pressure oil discharged by the second pump device In a hydraulic drive system for a construction machine having a control unit, provided in the second pump device, setting is changed according to the volume of the second pump device to generate a pressure according to the volume of the second pump device A variable valve device, a first detection device detecting a discharge pressure of the second pump device, a second detection device detecting a pressure generated by the variable valve device, a proportional solenoid valve, and an output of the proportional solenoid valve The pressure is introduced, and it is detected by the torque control piston which decreases the capacity of the first pump device by the increase of the output pressure, the discharge pressure of the second pump device detected by the first detection device, and the second detection device A controller for calculating an absorption torque of the second pump device from the output pressure of the variable valve device and outputting a drive signal corresponding to the absorption torque of the second pump device to the proportional solenoid valve; Example A solenoid valve operates according to the drive signal output from the control device to output a pressure according to the absorption torque of the second pump device to the torque control piston, and the torque control piston is the second pump The first maximum torque is reduced by the absorption torque of the second pump device by controlling the capacity of the first pump device by the pressure according to the absorption torque of the device.

  In the present invention configured as described above, the control device calculates the absorption torque of the second pump device based on the discharge pressure of the first pump device detected by the first and second detection devices and the generated pressure of the variable valve device. , The corresponding drive signal is output to a proportional solenoid valve. The proportional solenoid valve outputs a pressure corresponding to the absorption torque of the second pump device calculated by the control device to the torque control piston. As a result, when the absorption torque of the second pump device increases and reaches the second maximum torque, the torque control piston decreases the first maximum torque of the first pump device by the absorption torque of the second pump device. Also, even when the absorption torque of the second pump device is smaller than the second maximum torque, the torque control piston reduces the first maximum torque of the first pump device by the absorption torque of the second pump device. In the combined operation of simultaneously driving a plurality of actuators related to the first and second pump devices by this, it goes without saying that the second pump device is limited in torque control and is in an operating state operated at the second maximum torque. The first maximum torque of the first pump device is equal to the absorption torque of the second pump device even when the second pump device is in an operation state where the capacity control is performed by load sensing control without being limited by the torque control. It is corrected to be reduced, the total torque control can be accurately performed, and the rated output torque of the prime mover can be used effectively.

  In addition, a variable valve device is used to generate a pressure corresponding to the capacity of the second pump device, and the control device is used to absorb the absorption torque of the second pump device based on the discharge pressure of the second pump device and the generated pressure of the variable valve device. Since the pressure according to the calculated and calculated absorption torque is output from the proportional solenoid valve to the torque control piston, the circuit configuration is not complicated, and the pump device portion including the first and second pump devices is miniaturized. can do. Therefore, even with a small hydraulic shovel (in particular, a small rear swing type with a small rear end radius), the first and second pump devices can be easily mounted in a narrow space.

  According to the present invention, one of the hydraulic pumps has a pump control device that performs at least torque control, and the other hydraulic pump has at least two variable displacement hydraulic pumps that perform load sensing control and torque control. In the hydraulic drive system, the absorption torque of the other hydraulic pump is accurately detected with a simple configuration and fed back to one hydraulic pump side, thereby performing all torque control with high accuracy and effectively utilizing the rated output torque of the prime mover be able to.

It is a figure showing a hydraulic drive of a hydraulic shovel (construction machine) concerning a 1 embodiment of the present invention. It is a figure which shows the external appearance of the hydraulic shovel by which the hydraulic drive in this Embodiment is mounted. It is a figure which shows the pressure-reduction characteristic {The volume (tilt angle) of a main pump, and the relationship between the output pressure of a variable pressure-reduction valve} of a variable pressure-reduction valve. It is a block diagram which shows the calculation content of a controller. It is a figure which shows the torque control characteristic obtained by a 1st torque control part, and the effect of this Embodiment. It is a figure which shows the torque control characteristic obtained by a 2nd torque control part.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

~Constitution~
FIG. 1 is a view showing a hydraulic drive system of a hydraulic shovel (construction machine) according to an embodiment of the present invention.

  In FIG. 1, the hydraulic drive system according to the present embodiment is driven by a prime mover (for example, a diesel engine) 1 and the prime mover 1 to discharge pressurized oil to the first and second pressurized oil supply paths 105 and 205. And a split flow type variable displacement main pump 102 (first pump device) having a second discharge port 102a and 102b, and a third discharge driven by the prime mover 1 and discharging pressure oil to a third pressure oil supply passage 305 It is discharged from a single flow type variable displacement main pump 202 (second pump device) having a port 202a, first and second discharge ports 102a and 102b of the main pump 102, and a third discharge port 202a of the main pump 202, A plurality of acts driven by pressure oil supplied via the first to third pressure oil supply paths 105, 205, 305 The first and second discharge ports 102a and 102b of the main pump 102 are connected to the eta 3a, 3b, 3c, 3d, 3e, 3f, 3g and 3h and the first to third pressure oil supply paths 105, 205 and 305, respectively. And control valve units 104, 204, 304 for controlling the flow of hydraulic fluid supplied from the third discharge port 202a of the main pump 202 to the plurality of actuators 3a to 3h, and between the control valve unit 104 and the control valve unit 204. The discharge flow rate of the intermediate block 59 located, the regulator 112 (first pump control device) for controlling the discharge flow rate of the first and second discharge ports 102a and 102b of the main pump 102, and the third discharge port 202a of the main pump 202 Regulator 212 (the second pump control device) that controls the Eteiru.

  The control valve units 104, 204 and 304 are connected to the first to third pressure oil supply paths 105, 205 and 305, and the first and second discharge ports 102 a and 102 b of the main pump 102 and the third discharge of the main pump 202. A plurality of flow control valves 6a, 6b, 6c, 6d, 6e, 6f, 6g, 6h, 6i, 6j for controlling the flow of pressure oil supplied from the port 202a to the plurality of actuators 3a to 3h, and a plurality of flow control A plurality of pressure compensation valves 7a, 7b, 7c, 7d, 7e, 7f, 7g for controlling the differential pressure of the flow control valves 6a-6j respectively so that the differential pressure of the valves 6a-6j becomes equal to the target differential pressure. , 7h, 7i, 7j and a plurality of flow control valves 6a to 6j together with a plurality of operation detection valves 8a for detecting the switching of each flow control valve. 8b, 8c, 8d, 8e, 8f, 8g, 8h, 8i, 8j and the first pressure oil supply passage 105, which controls the pressure of the first pressure oil supply passage 105 not to exceed the set pressure. It is connected to the relief valve 114 and the second pressure oil supply passage 205, and is connected to the main relief valve 214 that controls the pressure of the second pressure oil supply passage 205 not to exceed the set pressure, and is connected to the third pressure oil supply passage 305 Is connected to the first pressure oil supply passage 105 so as to control the pressure of the third pressure oil supply passage 305 not to exceed the set pressure, and the pressure of the first pressure oil supply passage 105 is the first It becomes an open state when the pressure (unload valve set pressure) is the sum of the maximum load pressure of the actuator driven by the pressure oil discharged from the discharge port 102a and the set pressure (predetermined pressure) of the spring. A pressure oil connected to the unloading valve 115 for returning the pressure oil in the pressure oil supply passage 105 to the tank, and the second pressure oil supply passage 205, and the pressure in the second pressure oil supply passage 205 being discharged from the second discharge port 102b. Becomes higher than the pressure (unload valve set pressure) obtained by adding the set pressure (predetermined pressure) of the spring to the maximum load pressure of the actuator driven by the valve, and the pressure oil in the second pressure oil supply passage 205 is opened. The pressure in the third pressure oil supply passage 305 is set to the maximum load pressure of the actuator driven by the pressure oil discharged from the third discharge port 202a. It has an unload valve 315 that returns to the tank the pressure oil of the third pressure oil supply passage 305 when it becomes higher than the pressure (unload valve set pressure) obtained by adding the set pressure (predetermined pressure) of the spring. ing.

  The control valve units 104, 204, 304 are also connected to load ports of flow control valves 6c, 6d, 6f, 6i, 6j connected to the first pressure oil supply passage 105, and actuators 3a, 3b, 3c, 3d. , 3f, and the first load pressure detection circuit 131 including shuttle valves 9c, 9d, 9f, 9i, 9j for detecting the maximum load pressure Plmax 1 and the flow control valves 6b, 6e, connected to the second pressure oil supply passage 205. Second load pressure detection circuit 132 including shuttle valves 9b, 9e, 9g, 9h connected to load ports 6g, 6h and detecting maximum load pressure Plmax 2 of actuators 3b, 3e, 3g, 3h, and third pressure oil A third load pressure detection circuit 133 connected to the load port of the flow control valve 6a connected to the supply path 305 and detecting a load pressure (maximum load pressure) Plmax3 of the actuator 3a; A differential pressure which outputs a differential pressure (LS differential pressure) between the pressure (discharge pressure of the first discharge port 102a) P1 of the supply path 105 and the maximum load pressure Plmax1 detected by the first load pressure detection circuit 131 as an absolute pressure Pls1. Differential pressure (LS differential pressure) between the pressure reducing valve 111, the pressure of the second pressure oil supply passage 205 (discharge pressure of the second discharge port 102b) P2 and the maximum load pressure Plmax2 detected by the second load pressure detection circuit 132 Is output as the absolute pressure Pls2, the pressure of the third pressure oil supply path 305 (the discharge pressure of the third discharge port 202a) P3 and the maximum load pressure Plmax3 detected by the third load pressure detection circuit 133 Differential pressure reduction that outputs a differential pressure (LS differential pressure) with (the load pressure of the actuator 3a connected to the third pressure oil supply path 305-the load pressure of the boom cylinder 3a in the illustrated embodiment) as the absolute pressure Pls3 And a valve 311.

  The highest load pressure Plmax1 detected by the first load pressure detection circuit 131 is introduced to the above-described unload valve 115 as the highest load pressure of the actuator driven by the pressure oil discharged from the first discharge port 102a. The highest load pressure Plmax2 detected by the second load pressure detection circuit 132 is introduced to the unloaded valve 215 as the highest load pressure of the actuator driven by the pressure oil discharged from the second discharge port 102b, as described above. The maximum load pressure Plmax3 detected by the third load pressure detection circuit 133 is introduced to the unload valve 315 as the maximum load pressure of the actuator driven by the pressure oil discharged from the third discharge port 202a.

  Further, the LS differential pressure (absolute pressure Pls1) output by the differential pressure reducing valve 111 is the pressure compensating valves 7c, 7d, 7f, 7i, 7j connected to the first pressure oil supply passage 105 and the regulator 112 of the main pump 102. And the LS differential pressure (absolute pressure Pls2) output from the differential pressure reducing valve 211 is controlled by the pressure compensating valves 7b, 7e, 7g, 7h connected to the second pressure oil supply passage 205 and the regulator 112 of the main pump 102. The LS differential pressure (absolute pressure Pls3) output from the differential pressure reducing valve 311 is introduced to the pressure compensating valve 7a connected to the third pressure oil supply path 305 and the regulator 212 of the main pump 202.

  The actuator 3a is connected to the first discharge port 102a through the flow control valve 6i, the pressure compensation valve 7i and the first pressure oil supply passage 105, and the flow control valve 6a, the pressure compensation valve 7a and the third pressure oil supply passage It is connected to the third discharge port 202 a via 305. The actuator 3a is, for example, a boom cylinder for driving a boom of a hydraulic shovel, the flow control valve 6a is for main drive of the boom cylinder 3a, and the flow control valve 6i is for assist drive of the boom cylinder 3a. The actuator 3b is connected to the first discharge port 102a through the flow control valve 6j and the pressure compensation valve 7j and the first pressure oil supply passage 105, and the flow control valve 6b and the pressure compensation valve 7b and the second pressure oil supply passage It is connected to the second discharge port 102 b through 205. The actuator 3b is, for example, an arm cylinder that drives an arm of a hydraulic shovel, the flow control valve 6b is for main driving of the arm cylinder 3b, and the flow control valve 6j is for assist driving of the arm cylinder 3b.

  The actuators 3c, 3d, 3f are connected to the first discharge port 102a through the flow control valves 6c, 6d, 6f, the pressure compensation valves 7c, 7d, 7f and the first pressure oil supply passage 105, respectively, and the actuators 3g, 3e, 3h is connected to the second discharge port 102b through the flow control valves 6g, 6e, 6h, the pressure compensating valves 7g, 7e, 7h and the second pressure oil supply passage 205, respectively. The actuators 3c, 3d, and 3f are, for example, a swing motor that drives an upper swing body of a hydraulic shovel, a bucket cylinder that drives a bucket, and a left travel motor that drives a left crawler belt of a lower travel body. The actuators 3g, 3e and 3h are, for example, a right traveling motor that drives a right crawler belt of a lower traveling body of a hydraulic shovel, a swing cylinder that drives a swing post, and a blade cylinder that drives a blade.

  The control valve units 104, 204, 304 and the intermediate block 59 are connected such that the upstream side is connected to the pilot pressure oil supply path 31b (described later) via the throttle 43 and the downstream side is connected to the tank via the operation detection valves 8a to 8j. A combined operation detection oil passage 53 and a first switching valve 40, a second switching valve 146 and a third switching valve 246 which are switched based on the operation detection pressure generated by the traveling combined operation detection oil passage 53 are provided. .

  The traveling combined operation detection oil passage 53 is at least one of the operation detecting valves 8a to 8j when the traveling combined operation for simultaneously driving the left traveling motor 3f and / or the right traveling motor 3g and at least one other actuator is not performed. The pressure in the oil passage becomes the tank pressure by communicating with the tank via the valve, and at the time of combined travel operation, either of the operation detection valves 8f and 8g and the operation detection valves 8a to 8j together with the corresponding flow control valve And the communication with the tank is interrupted to generate an operation detection pressure (operation detection signal).

  When the first switching valve 40 is not in the traveling combined operation, it is in the first position (shutdown position) on the lower side in the drawing and blocks the communication between the first pressure oil supply passage 105 and the second pressure oil supply passage 205; At the time of travel complex operation, the first pressure oil supply path 105 and the second pressure oil supply path 205 are switched to the second position (communication position) on the upper side of the figure by the operation detection pressure generated in the travel complex operation detection oil path 53. Communicate.

  The second switching valve 146 is at the first position on the lower side in the drawing when it is not the traveling combined operation, and guides the tank pressure to the shuttle valve 9h on the most downstream side of the second load pressure detection circuit 132. The maximum load pressure Plmax1 detected by the first load pressure detection circuit 131 is switched to the second position in the upper side of the drawing by the operation detection pressure generated in the traveling combined operation detection oil passage 53 of the second load pressure detection circuit 132. It leads to the most downstream shuttle valve 9h.

  The third switching valve 246 is at the first position on the lower side in the drawing when it is not traveling combined operation, and guides the tank pressure to the most downstream shuttle valve 9i of the first load pressure detection circuit 131, and at the time of traveling combined operation The maximum load pressure Plmax2 detected by the second load pressure detection circuit 132 is switched to the second position in the upper side of the drawing by the operation detection pressure generated in the traveling combined operation detection oil passage 53 of the first load pressure detection circuit 131. It leads to the most downstream shuttle valve 9i.

  The hydraulic drive system according to the present embodiment is connected to the fixed displacement pilot pump 30 driven by the prime mover 1 and the pressure oil supply passage 31a of the pilot pump 30, so that the discharge flow rate of the pilot pump 30 is absolute pressure Pgr. And a pilot relief valve 32 connected to a pilot pressure oil supply passage 31b downstream of the prime mover rotation number detection valve 13 and generating a constant pilot pressure in the pilot pressure oil supply passage 31b. A gate lock valve 100 which is connected to the pilot pressure oil supply path 31b and switches whether the pilot pressure oil supply path 31c on the downstream side is connected to the pilot pressure oil supply path 31b or to the tank by the gate lock lever 24; It is connected to the pilot pressure oil supply passage 31c on the downstream side of the lock valve 100, and a plurality of flow control A plurality of operating devices 122, 123, 124a, 124b having a plurality of pilot valves (pressure reducing valves) for generating operation pilot pressure for controlling the valves 6a, 6b, 6c, 6d, 6e, 6f, 6g, 6h (figure 2) and.

  The prime mover rotational speed detection valve 13 includes a flow rate detection valve 50 connected between the pressure oil supply path 31a of the pilot pump 30 and the pilot pressure oil supply path 31b, and a differential pressure across the flow rate detection valve 50 as an absolute pressure Pgr. And a differential pressure reducing valve 51 for outputting as

  The flow rate detection valve 50 has a variable throttling portion 50a that increases the opening area as the passing flow rate (discharge flow rate of the pilot pump 30) increases. The discharge oil of the pilot pump 30 passes through the variable throttle portion 50a of the flow rate detection valve 50 and flows toward the pilot pressure oil supply path 31b. At this time, a differential pressure that increases as the passing flow rate increases is generated in the variable throttle portion 50a of the flow rate detection valve 50, and the differential pressure reducing valve 51 outputs the differential pressure before and after as the absolute pressure Pgr. Since the discharge flow rate of the pilot pump 30 changes according to the rotational speed of the prime mover 1, the discharge flow rate of the pilot pump 30 can be detected by detecting the differential pressure across the variable throttle portion 50a. It can be detected.

  The regulator 112 of the main pump 102 selects the low pressure side of the LS differential pressure (absolute pressure Pls1) output by the differential pressure reducing valve 111 and the LS differential pressure (absolute pressure Pls2) output by the differential pressure reducing valve 211 The LS control valve 112b operates by the differential pressure between the LS differential pressure 112a and the low pressure selected LS differential pressure and the output pressure (absolute pressure) Pgr of the motor rotation speed detection valve 13, and when the LS differential pressure> Pgr, the input side Is communicated with the pilot pressure oil supply passage 31b to raise the output pressure, and when LS differential pressure <Pgr, the LS control valve 112b which causes the input side to communicate with the tank to reduce the output pressure, and the output of the LS control valve 112b The pressure is introduced, and the pressure (the volume) of the main pump 102 is reduced by the increase in the output pressure, and the pressures of the first and second pressure oil supply paths 105 and 205 of the main pump 102 are reduced. Led, The torque control (horsepower control) pistons 112e and 112d decrease the displacement (capacity) of the main pump 102 by the increase in pressure of the main pump 102, and the pressure corresponding to the absorption torque of the main pump 202 is derived. A torque control (horsepower control) piston 112f for reducing the displacement (capacity) of the pump 102 and a spring 112u for setting a maximum torque T12max (see FIG. 5A) available to the main pump 102 are provided.

  The regulator 212 of the main pump 202 is operated by a differential pressure between the LS differential pressure (absolute pressure Pls3) output from the differential pressure reducing valve 311 and the output pressure (absolute pressure) Pgr of the motor rotational speed detection valve 13 When LS differential pressure> Pgr, the input side is communicated with pilot pressure oil supply path 31b to increase the output pressure, and when LS differential pressure <Pgr, the input side is communicated with the tank and output The LS control valve 212b reduces the pressure, and the LS control piston 212c reduces the displacement (capacity) of the main pump 202 by increasing the output pressure of the LS control valve 212b. The pressure of the 3-pressure oil supply passage 305 is introduced, and a torque control (horsepower control) piston 212 d that reduces the displacement (capacity) of the main pump 202 by the increase of the pressure is used. And a spring 212e that sets the maximum torque T3max (see FIG. 5B) that can be used.

  The low pressure selection valve 112a, the LS control valve 112b, and the LS control piston 112c of the regulator 112 are pressure oil from which the discharge pressure of the first and second discharge ports 102a and 102b is discharged from the first and second discharge ports 102a and 102b. The first load sensing control unit is configured to control the displacement of the main pump 102 so as to be higher than the maximum load pressure of the actuator driven by the target differential pressure. The LS control valve 212b and the LS control piston 212c of the regulator 212 have the discharge pressure of the third discharge port 202a higher by the target differential pressure than the maximum load pressure of the actuator driven by the pressure oil discharged from the third discharge port 202a. The second load sensing control unit configured to control the displacement of the main pump 202 is configured.

  The torque control pistons 112d and 112e, the torque control piston 112f, and the spring 112u of the regulator 112 increase the average pressure of the discharge pressure of the first discharge port 102a and the discharge pressure of the second discharge port 102b. A first torque control unit is configured to reduce the displacement and reduce the displacement of the main pump 102 (first pump device) as the discharge pressure of the third discharge port 202a increases, and the torque control piston 212d of the regulator 212 The spring 212e constitutes a second torque control unit that reduces the displacement of the main pump 202 as the discharge pressure of the third discharge port 202a increases.

  FIG. 2 is a view showing the appearance of a hydraulic shovel on which the above-described hydraulic drive system is mounted.

  In FIG. 2, the hydraulic shovel well-known as a work machine includes a lower traveling body 101, an upper swing body 150, and a swing type front work machine 151, and the front work machine 151 includes a boom 152, an arm 153, It is composed of a bucket 154. The upper swing body 150 can be pivoted by the swing motor 3 c with respect to the lower traveling body 101. A swing post 103 is attached to a front portion of the upper swing body 150, and a front working machine 151 is attached to the swing post 103 so as to be vertically movable. The swing post 103 is rotatable horizontally with respect to the upper swing body 150 by the expansion and contraction of the swing cylinder 3e, and the boom 152, the arm 153 and the bucket 154 of the front working machine 151 are the boom cylinder 3a, the arm cylinder 3b, and the bucket cylinder It can be vertically rotated by the expansion and contraction of 3d. The central frame of the lower traveling body 101 is attached with a blade 106 that moves up and down by extension and contraction of the blade cylinder 3 h. The lower traveling body 101 travels by driving the left and right crawler belts 101a and 101b by the rotation of the traveling motors 3f and 3g.

  A driver's cab 108 is installed in the upper revolving superstructure 150, and in the driver's cab 108, a driver's seat 121, front / right operation devices 122 and 123 for turning (only illustrated on the left side in FIG. 3), operation devices for traveling An operation device for swing and an operation device for a blade, a gate lock lever 24 and the like (not shown) are provided. The operation levers of the operation devices 122 and 123 can be operated from the neutral position in any direction based on the cross direction, and when operating the operation levers of the left operation device 122 in the front and back direction, the operation device 122 is for turning When operating as an operating device and operating the operating lever of the operating device 122 in the left and right direction, the operating device 122 functions as an operating device for the arm, and when operating the operating lever of the right operating device 123 in the front and rear direction, The operating device 123 functions as an operating device for a boom, and when operating the operating lever of the operating device 123 in the left-right direction, the operating device 123 functions as an operating device for a bucket.

  Returning to FIG. 1, the hydraulic drive system according to the present embodiment includes a proportional solenoid valve 61, a variable pressure reducing valve 62 (variable valve device), a pressure sensor 63 (first detection device), a pressure sensor 64 (second detection device) The controller 65 (control device) is further provided.

  The proportional solenoid valve 61 has an input side connected to the pilot pressure oil supply path 31b and the tank, and an output side connected to the torque control piston 112f, and operates according to a drive signal output from the controller 65 to operate the pressure of the pilot pressure oil supply path 31b. And generates a pressure according to the absorption torque of the main pump 202, and outputs the pressure to the torque control piston 112f.

  The variable pressure reducing valve 62 has an input side connected to the pilot pressure oil supply path 31b and the tank, and an output side connected to the oil path 66, and the pressure of the pilot pressure oil supply path 31b according to the volume (tilt angle) of the main pump 202 Is reduced, and the reduced pressure is output to the oil passage 66.

  The pressure sensor 63 is connected to the third pressure oil supply passage 305 of the main pump 202, and detects the pressure of the third pressure oil supply passage 305 (the discharge pressure of the main pump 202). The pressure sensor 64 is connected to the oil passage 66 and detects the output pressure of the variable pressure reducing valve 62.

  The controller 65 receives detection signals from the pressure sensors 63, 64, and calculates the absorption torque of the main pump 202 based on the discharge pressure of the main pump 202 detected by the pressure sensors 63, 64 and the output pressure of the variable pressure reducing valve 62, The corresponding drive signal is output to the proportional solenoid valve 61. The proportional solenoid valve 61 operates according to a drive signal output from the controller 65, and outputs a pressure corresponding to the absorption torque of the main pump 202 calculated by the controller 65 to the torque control piston 112f.

  The main pumps 102 and 202 are, for example, swash plate pumps provided with variable displacement members 102d and 202d, respectively, and change pump displacements when the tilt angles of the variable displacement members 102d and 202d change.

  The variable pressure reducing valve 62 has a setting spring 62 a that biases in the pressure reducing direction, and a piston 62 b that changes the biasing force (spring force) of the setting spring 62 a. The piston 62 b corresponds to the variable displacement member 202 d of the main pump 202 It abuts and is displaced according to the tilt angle of the variable capacity member 202d to change the biasing force of the setting spring 62a. The variable pressure reducing valve 62 reduces the pressure in the pilot pressure oil supply passage 31b in accordance with the change in the biasing force of the set spring 62a.

  FIG. 3 is a diagram showing the pressure reduction characteristics of the variable pressure reducing valve 62 (the relationship between the volume (tilt angle) of the main pump 202 and the output pressure of the variable pressure reducing valve 62).

  When the tilt angle of the variable displacement member 202d of the main pump 102 is minimum, the displacement of the piston 52b is minimum, and the biasing force of the setting spring 62a is also minimum. At this time, the output pressure of the variable pressure reducing valve 62 is maximum. As the tilt angle of the variable capacity member 202d increases, the displacement of the piston 52b also increases, and the biasing force of the setting spring 62a also increases. As a result, the variable pressure reducing valve 62 reduces the pressure in the pilot pressure oil supply passage 31b according to the biasing force of the setting spring 62a, and the output pressure of the variable pressure reducing valve 62 is the tilt angle of the variable displacement member 202d as shown in FIG. Changes as it increases. When the tilt angle of the variable displacement member 202d of the main pump 102 is maximized, the displacement of the piston 52b is also maximized, and the biasing force of the setting spring 62a is also maximized. Thus, the output pressure of the variable pressure reducing valve 62 is minimized.

  FIG. 4 is a block diagram showing the calculation content of the controller 65. As shown in FIG.

  The controller 65 includes a pump displacement conversion block 65a, a torque calculation block 65b, a pressure conversion block 65c, and a current conversion block 65d.

  The pump displacement conversion block 65 a converts the output pressure of the variable pressure reducing valve 62 detected by the pressure sensor 63 into the displacement of the main pump 202 using table data preset in the controller 65. The table data has an inverse characteristic of the pressure reducing characteristic of the variable pressure reducing valve 62 shown in FIG.

  The absorption torque of the hydraulic pump can be obtained by the following equation.

(Discharge pressure × pump capacity) / 2π
The torque calculation block 65b calculates the absorption torque of the main pump 202 from the discharge pressure of the main pump 202 detected by the pressure sensor 63 and the capacity of the main pump 202 obtained by the pump displacement conversion block 65a according to the above equation.

  The pressure conversion block 65c uses table data preset in the controller 65, and the output pressure of the proportional solenoid valve 61 necessary for performing the torque reduction control of the main pump 102 by the absorption torque determined by the torque calculation block 65b (main The absorbed torque of the pump 202 is converted into the pressure required for torque reduction control by the torque control piston 112f.

  The current conversion block 65d converts the pressure determined by the pressure conversion block 65c into a current value necessary for the proportional solenoid valve 61 to output, and the controller 65 outputs the current value as a drive signal to the proportional solenoid valve 61.

  The output pressure of the proportional solenoid valve 61 is guided to the torque control piston 112f of the main pump 102, and the torque control piston 112f is driven by the output pressure of the proportional solenoid valve 61 (pressure according to the absorption torque of the main pump 202). Is controlled to reduce the maximum torque T12max available to the main pump 102 by the absorption torque of the main pump 202.

  FIG. 5A is a view showing the torque control characteristic obtained by the first torque control unit (torque control pistons 112 d and 112 e, the torque control piston 112 f, and the spring 112 u) and the effect of the present embodiment. In FIG. 5A, P12 is the sum P1 + P2 of the discharge pressures P1 and P2 of the first and second discharge ports 102a and 102b of the main pump 102 (discharge pressure of the main pump 102), and q12 is the swash plate of the main pump 102. P12max is the sum of the maximum discharge pressures of the first and second discharge ports 102a and 102b of the main pump 102 obtained by the main relief valve 114 and 214, and q12max is the structure of the main pump 102 Is the maximum tilt angle determined by The absorption torque of the main pump 102 is represented by the product of the discharge pressure P12 (P1 + P2) of the main pump 102 and the tilting angle q12.

  In FIG. 5A, the maximum torque available to the main pump 102 is set by the spring 112 u to T 12 max shown by the curve 502. The actuator is driven by the pressure oil discharged from the main pump 102, and when the absorption torque of the main pump 102 increases and reaches the maximum torque T12max, the main pump 102 does not increase its absorption torque any more. The angle is limited by the torque control pistons 112d, 112e of the regulator 112. For example, when the discharge pressure of the main pump 102 increases while the tilt angle of the main pump 102 is on any of the curves 502, the torque control pistons 112d and 112e follow the curve 502 with the tilt angle q12 of the main pump 102. Reduce. When the tilt angle q12 of the main pump 102 is to be increased while the tilt angle of the main pump 102 is on any of the curves 502, the torque control pistons 112d and 112e have the tilt angle q12 of the main pump 102. The tilt angle on the curve 502 is restricted to be held. In FIG. 5A, a symbol TE is a curve showing the rated output torque Terate of the prime mover 1, and the maximum torque T12max is set to a value smaller than the Terate. By thus setting the maximum torque T12max and limiting the absorption torque of the main pump 102 so as not to exceed the maximum torque T12max, the main pump 102 can utilize the rated output torque Terate of the prime mover 1 as effectively as possible. A stop (engine stall) of the prime mover 1 when driving the actuator can be prevented.

  FIG. 5B is a view showing a torque control characteristic obtained by the second torque control unit (the torque control piston 212 d and the spring 212 e). In FIG. 5B, P3 is the discharge pressure of the main pump 202, q3 is the tilt angle (capacity) of the swash plate of the main pump 202, and P3max is the maximum discharge pressure of the main pump 202 obtained by the main relief valve 314. And q3max is the maximum tilt angle determined by the structure of the main pump 202. The absorption torque of the main pump 202 is represented by the product of the discharge pressure P3 of the main pump 202 and the tilt angle q3.

  In FIG. 5B, the maximum torque available to the main pump 202 is set by the spring 212e to T3max shown by a curve 602. When the actuator is driven by the pressure oil discharged from the main pump 202 and the absorption torque of the main pump 202 increases and reaches the maximum torque T3max, the absorption torque of the main pump 202 is equal to that of the regulator 112 of FIG. 5A. The tilt angle of the main pump 202 is limited by the torque control piston 212 d of the regulator 212 so as not to increase more than that.

  Arrows AR1 and AR2 in FIG. 5A indicate the effects of the torque control piston 112f. When the discharge pressure P3 or the capacity of the main pump 202 increases and the absorption torque increases, the torque control piston 112f performs the main pump the maximum torque T12max set by the spring 112u as shown by arrows AR1 and AR2 in FIG. 5A. Decrease the absorption torque of 202. As a result, even in the combined operation of simultaneously driving the actuator related to the main pump 102 and the actuator related to the main pump 202, the absorption torque of the main pump 102 is controlled so as not to exceed the maximum torque T12max (total torque control). Stop (engine stall) can be prevented. In FIG. 5A, arrow AR1 indicates that the main pump 202 (second hydraulic pump) is subject to torque control limitation and operates at the maximum torque control T3max, and arrow AR2 indicates that the main pump 202 is torque controlled. In the case of performing capacity control by load sensing control without being subject to the limitation of

~ Operation ~
Next, the operation of the present embodiment will be described.

(A) When all the control levers are neutral Since the control levers of all the control devices are neutral, all the flow control valves 6a to 6j are in the neutral position. Since all of the flow control valves 6a to 6j are in the neutral position, the first load pressure detection circuit 131, the second load pressure detection circuit 132, and the third load pressure detection circuit 133 respectively have the maximum load pressure Plmax1, Plmax2 and Plmax3 as the tank. Detect pressure. The maximum load pressures Plmax1, Plmax2 and Plmax3 are led to the unloading valves 115, 215 and 315 and the differential pressure reducing valves 111, 211 and 311, respectively. The unloading valves 115, 215, and 315 have respective pressures P1, P2, and P3 of the first, second, and third discharge ports 102a, 102b, and 202a at the tank pressure, and the springs of the unloading valves 115, 215, and 315, respectively. When the pressure becomes higher than the sum of the set pressure, the pressure oil is returned to the tank. Assuming that the set pressure of the spring of the unload valve 115, 215, 315 is Punsp, Punsp is set slightly higher than the output pressure Pgr of the motor rotation speed detection valve 13 which is the target LS differential pressure (Punsp> Pgr).

  The differential pressure reducing valves 111, 211 and 311 are respectively pressure P1, P2 and P3 of the first, second and third pressure oil supply paths 105, 205 and 305 and maximum load pressures Plmax1, Plmax2 and Plmax3 (tank pressure). And the differential pressure (LS differential pressure) with each other are output as absolute pressures Pls1, Pls2, and Pls3. Since the maximum load pressures Plmax1, Plmax2 and Plmax3 are tank pressures as described above, Pls1 = Pls2 = Pls3 ≒ Punsp> Pgr. The LS differential pressure Pls 1, Pls 2 is led to the low pressure selection valve 112 a of the regulator 112, and Pls 3 is led to the LS control valve 212 b of the regulator 212.

  In the regulator 112, the low pressure side of the LS differential pressures Pls1 and Pls2 led to the low pressure selection valve 112a is selected and led to the LS control valve 112b. At this time, even if Pls1 or Pls2 is selected, Pls1 or Pls2> Pgr, so the LS control valve 122b is pushed leftward in the figure and switched to the right position, and generated by the pilot relief valve 32. Constant pilot pressure to the LS control piston 112c. Since the pilot pressure is introduced to the LS control piston 112c, the displacement of the main pump 102 is kept to a minimum.

  On the other hand, the LS differential pressure Pls3 is introduced to the LS control valve 212b of the regulator 212. Since Pls3> Pgr, the LS control valve 212b is pushed in the right direction in the figure to switch to the left position, and guides a constant pilot pressure generated by the pilot relief valve 32 to the LS control piston 212c. Since the pilot pressure is introduced to the LS control piston 212c, the displacement of the main pump 202 is kept to a minimum.

(B) Single operation of boom For example, when the boom control lever is fully operated in the direction in which the boom cylinder 3a extends, that is, in the boom raising direction, the flow control valves 6a and 6i for driving the boom cylinder 3a move upward in the figure. The opening area of the meter-in passage of the flow control valves 6a and 6i is maximized.

  The load pressure on the bottom side of the boom cylinder 3 a is detected by the third load pressure detection circuit 133 as the maximum load pressure Plmax 3, and is led to the unload valve 315 and the differential pressure reducing valve 311. By the highest load pressure Plmax3 being led to the unloading valve 315, the unloading valve 315 shuts off the oil passage for discharging the pressure oil of the third pressure oil supply passage 305 to the tank. Further, the maximum load pressure Plmax3 is introduced to the differential pressure reducing valve 311, so that the differential pressure reducing valve 311 absolutely measures the differential pressure (LS differential pressure) between the pressure P3 of the third pressure oil supply passage 305 and the maximum load pressure Plmax3. Output as pressure Pls3. This Pls3 is led to the LS control valve 212b. The LS control valve 212 b compares the output pressure Pgr of the motor rotation speed detection valve 13 which is the target LS differential pressure with the above Pls 3.

Since the pressure in the third pressure oil supply passage 305 and the load pressure in the boom cylinder 3a are almost equal and Pls3, which is the LS differential pressure, is almost 0 immediately after Pls 3 <Pgr, immediately after the operation lever input at the boom raising start, The control valve 212b switches to the left in the drawing and discharges the pressure oil of the LS control piston 212c to the tank. Therefore, the displacement of the main pump 202 increases, and the discharge flow rate of the main pump 202 increases accordingly, and the flow increase continues until Pls3 = Pgr. As a result, pressure oil at a flow rate corresponding to the input of the boom control lever is supplied to the bottom side of the boom cylinder 3a, and the boom cylinder 3a is driven in the extension direction.

  On the other hand, the load pressure on the bottom side of the boom cylinder 3 a is detected by the first load pressure detection circuit 131 as the maximum load pressure Plmax 1 and is led to the unload valve 115 and the differential pressure reducing valve 111. As the maximum load pressure Plmax1 is introduced to the unloading valve 115, the unloading valve 115 shuts off the oil passage for discharging the pressure oil of the first pressure oil supply passage 105 to the tank. Further, the maximum load pressure Plmax1 is introduced to the differential pressure reducing valve 111, so that the differential pressure reducing valve 111 absolutely measures the differential pressure (LS differential pressure) between the pressure P1 of the first pressure oil supply passage 105 and the maximum load pressure Plmax1. Output as pressure Pls1. This Pls1 is led to the low pressure selection valve 112a of the regulator 112, and the low pressure side of Pls1 and Pls2 is selected by the low pressure selection valve 112a.

Immediately after the boom raising start, the pressure of the first pressure oil supply passage 105 and the load pressure of the boom cylinder 3a are almost equal and Pls1, which is LS differential pressure, is almost 0, and Pls2 is the same as neutral of the operation lever. , Is maintained at a value greater than Pgr. Therefore, the low pressure selection valve 112a selects Pls1, and this Pls1 is led to the LS control valve 112b. The LS control valve 112b compares the output pressure Pgr of the motor rotation speed detection valve 13 which is the target LS differential pressure with Pls1, and since Pls1 <Pgr, the LS control valve 112b switches to the right in the figure. The pressure oil of the control piston 112c is discharged to the tank. Therefore, the displacement of the main pump 102 is increased, and the discharge flow rate of the main pump 202 is increased accordingly, and the flow increase is continued until Pls1 = Pgr. As a result, pressure oil at a flow rate corresponding to the input of the boom control lever is supplied from the first discharge port 102a of the main pump 102 to the bottom side of the boom cylinder 3a. The boom cylinder 3a is connected to the third discharge port 202a of the main pump 202. It is driven in the extension direction by the combined pressure oil from the first discharge port 102 a of the main pump 102.

  At this time, pressure oil having the same flow rate as the pressure oil supplied to the first pressure oil supply path 105 is supplied to the second pressure oil supply path 205, but the pressure oil is supplied as an excess flow rate via the unloading valve 215. Is returned to the tank.

  The controller 65 also calculates the absorption torque of the main pump 202 based on the discharge pressure of the main pump 202 detected by the pressure sensors 63 and 64 and the output pressure of the variable pressure reducing valve 62, and outputs a corresponding drive signal to the proportional solenoid valve 61. Output. The proportional solenoid valve 61 outputs a pressure according to the absorption torque of the main pump 202 calculated by the controller 65 to the torque control piston 112 f. As a result, when the absorption torque of the main pump 202 increases and reaches T3max in FIG. 5B, the torque control piston 112f generates the maximum torque T12max set by the spring 112u of the main pump 202, as shown by arrow AR1 in FIG. Reduce the absorption torque. Further, even when the absorption torque of the main pump 202 is T3g smaller than T3max in FIG. 5B, the torque control piston 112f performs the main pump with the maximum torque T12max set by the spring 112u as shown by the arrow AR2 in FIG. 5A. Decrease the absorption torque of 202. Thus, when the boom cylinder 3a is driven by the discharge flow rates of both the main pump 102 and the main pump 202, it goes without saying that the main pump 202 is in an operating state of being limited by the torque control and operating at the maximum torque T3max. Even if main pump 202 is not restricted by torque control and is in an operating state where capacity control is performed by load sensing control, correction is made such that maximum torque T12max of main pump 102 is reduced by the absorption torque of main pump 202. , It is possible to perform all torque control with high accuracy and to effectively use the rated output torque of the motor.

(C) Independent operation of arm For example, when the arm control lever is fully extended in the direction in which the arm cylinder 3b extends, that is, in the arm cloud direction, the flow control valves 6b and 6j for driving the arm cylinder 3b move downward in the figure. The opening area of the meter-in passage of the flow control valves 6b and 6j is maximized.

  The load pressure on the bottom side of the arm cylinder 3 b is detected by the first load pressure detection circuit 131 as the maximum load pressure Plmax 1 and is led to the unload valve 115 and the differential pressure reducing valve 111. As the maximum load pressure Plmax1 is introduced to the unloading valve 115, the unloading valve 115 shuts off the oil passage for discharging the pressure oil of the first pressure oil supply passage 105 to the tank. Further, the maximum load pressure Plmax1 is introduced to the differential pressure reducing valve 111, whereby the LS differential pressure Pls1 is introduced to the low pressure selection valve 112a of the regulator 112.

  On the other hand, the load pressure on the bottom side of the arm cylinder 3 b is detected by the second load pressure detection circuit 132 as the maximum load pressure Plmax 2 and is led to the unload valve 215 and the differential pressure reducing valve 211. The highest load pressure Plmax2 is introduced to the unloading valve 215, whereby the unloading valve 215 shuts off the oil passage for discharging the pressure oil of the second pressure oil supply passage 205 to the tank. Further, the maximum load pressure Plmax2 is introduced to the differential pressure reducing valve 211, whereby the LS differential pressure Pls2 is introduced to the low pressure selection valve 112a of the regulator 112.

Immediately after the control lever input at arm cloud start-up, the pressure of the first and second pressure oil supply paths 105, 205 is almost equal to the load pressure of the arm cylinder 3b, and both LS differential pressure Pls1, Pls2 is approximately 0 is there. Therefore, in the low pressure selection valve 112a, either Pls1 or Pls2 is selected as the low pressure side, and since Pls1 or Pls2 <Pgr, the LS control valve 112b switches to the right in the figure, and the pressure of the LS control piston 112c Release the oil to the tank. For this reason, the displacement of the main pump 102 increases, and the discharge flow rate of the main pump 202 increases accordingly, and the flow increase continues until Pls1 or Pls2 = Pgr. As a result, pressure oil at a flow rate corresponding to the input of the arm control lever is supplied from the first and second discharge ports 102a and 102b of the main pump 102 to the bottom side of the arm cylinder 3b, and the arm cylinder 3b receives the first and second discharges. It is driven in the extension direction by the combined pressure oil from the ports 102a and 102b.

  Also. Thus, when the arm cylinder 3b is driven by the discharge flow rates of the two discharge ports 102a and 102b of the main pump 102, the absorption torque of the main pump 102 is controlled so as not to exceed the maximum torque T12max in FIG. 5A. The engine stall can be prevented if the load on the engine 1 increases.

(D) Combined operation of boom and arm The combined operation of the boom and the arm is, for example, a combination of the boom independent operation and the arm independent operation described above. At this time, in the regulator 112 of the main pump 102, the LS control valve 12b operates with a balance between the low pressure side of LS differential pressure Pls1 and Pls2 and Pgr, so the load pressure of the boom cylinder 3a is higher than the load pressure of the arm cylinder 3b, When Pls1 <Pls2, pressure oil at a flow rate corresponding to the load pressure of the boom cylinder 3a is discharged from the first discharge port 102a of the main pump 102, and the pressure oil is supplied to the bottom of the boom cylinder 3a, and the arm cylinder 3b When the load pressure of the boom cylinder 3a is higher than the load pressure of the boom cylinder 3a, and Pls1> Pls2, pressure oil at a flow rate corresponding to the load pressure of the arm cylinder 3b is discharged from the first discharge port 102a of the main pump 102 Oil is supplied to the bottom side of the arm cylinder 3b.

  Thus, when the load pressure of the boom cylinder 3a is higher than the load pressure of the arm cylinder 3b, the boom cylinder 3a is a combined pressure oil from the third discharge port 202a of the main pump 202 and the first discharge port 102a of the main pump 102. The arm cylinder 3b is driven in the extension direction by the pressure oil from the second discharge port 102b of the main pump 102. When the load pressure of arm cylinder 3 b is higher than the load pressure of boom cylinder 3 a, boom cylinder 3 a is driven in the extending direction by the pressure oil from third discharge port 202 a of main pump 202, and arm cylinder 3 b is main pump 102. It is driven in the extension direction by the combined pressure oil from the first and second discharge ports 102a and 102b.

  The controller 65 also calculates the absorption torque of the main pump 202 based on the discharge pressure of the main pump 202 detected by the pressure sensors 63 and 64 and the output pressure of the variable pressure reducing valve 62, and outputs a corresponding drive signal to the proportional solenoid valve 61. Output. The proportional solenoid valve 61 outputs a pressure according to the absorption torque of the main pump 202 calculated by the controller 65 to the torque control piston 112 f. As a result, when the absorption torque of the main pump 202 increases and reaches T3max in FIG. 5B, the torque control piston 112f generates the maximum torque T12max set by the spring 112u of the main pump 202, as shown by arrow AR1 in FIG. Reduce the absorption torque. Further, even when the absorption torque of the main pump 202 is T3g smaller than T3max in FIG. 5B, the torque control piston 112f performs the main pump with the maximum torque T12max set by the spring 112u as shown by the arrow AR2 in FIG. 5A. Decrease the absorption torque of 202. As a result, when the boom cylinder 3a and the arm cylinder 3b are simultaneously driven, the main pump 202 is subjected to the torque control limitation, and the main pump 202 is limited to the torque control as a matter of course when operating under the maximum torque T3max. Is corrected so that the maximum torque T12max of the main pump 102 is reduced by the absorption torque of the main pump 202, and all torque control is performed with high accuracy. The rated output torque of the motor can be used effectively.

~effect~
As described above, according to the present embodiment, in the combined operation in which the main pump 102 and the plurality of actuators related to the main pump 202 are simultaneously driven, the main pump 202 is subjected to the torque control limitation and operates with the maximum torque T3max. Even when the main pump 202 is in an operating state in which the capacity control is performed by load sensing control without being limited by the torque control, as a matter of course when in the state, the maximum torque T12max of the main pump 102 is the main pump 202. It is corrected to reduce the absorption torque, the total torque control can be accurately performed, and the rated output torque of the prime mover can be effectively used.

  The variable pressure reducing valve 62 is used to generate a pressure according to the volume of the main pump 202, and the controller 65 calculates the absorption torque of the main pump 202 based on the discharge pressure of the main pump 202 and the generated pressure of the variable pressure reducing valve 62. The proportional solenoid valve 61 outputs a pressure corresponding to the calculated absorption torque to the torque control piston 112 f. As described above, in the present embodiment, since one proportional solenoid valve 61 is used and the tilt angle of the main pump 202 is hydraulically / electrically detected, the circuit configuration is not complicated, and the main pump 102, The pump device portion including 202 can be miniaturized. Therefore, even with a small hydraulic shovel (in particular, a small rear swing type with a small rear end radius), the pump device portion can be easily mounted in a narrow space.

  Furthermore, in the present embodiment, the variable pressure reducing valve 62 includes a setting spring 62a that biases the pressure reducing direction, and a piston 62b that changes the biasing force of the setting spring 62a. It abuts on the capacity member 202d and is displaced according to the tilt angle of the variable capacity member 202d to change the biasing force of the setting spring 62a, and the variable pressure reducing valve 62 changes the pilot pressure according to the variation of the biasing force of the setting spring 62a. The pressure of the oil supply passage 31 b is reduced and output. This eliminates the need for a dedicated auxiliary pump for hydraulically detecting the tilt angle of the hydraulic pump, which also simplifies the circuit configuration.

~ Other ~
Although the above embodiment has described the case where the construction machine is a hydraulic shovel, the first pump control device has the first torque control unit, and the second pump control device has the second torque control unit and the load sensing control. The present invention may be applied to a construction machine other than a hydraulic shovel, such as a wheel shovel, a hydraulic traveling crane, etc., as long as the construction machine is equipped with a hydraulic drive device having a part.

  Moreover, although the said embodiment demonstrated the case where a 1st pump apparatus was a hydraulic pump of a split flow type, a 1st pump apparatus may be a hydraulic pump of a single flow type.

  Moreover, the load sensing system of the said embodiment is an example, and a load sensing system can be variously deformed. For example, in the above embodiment, a differential pressure reducing valve that outputs the pump discharge pressure and the maximum load pressure as absolute pressure is provided, the output pressure is introduced to the pressure compensating valve, and the target compensating differential pressure is set. Although the target differential pressure of the load sensing control is set by guiding, the pump discharge pressure and the maximum load pressure may be guided to the pressure control valve or the LS control valve through separate oil passages.

1 Prime mover 3a to 3h Multiple actuators 6a to 6j Flow control valves 7a to 7j Pressure compensation valves 9c to 9j Shuttle valve 30 Pilot pump 61 Proportional solenoid valve 62 Variable pressure reducing valve (variable valve device)
62a setting spring 62b piston 63 pressure sensor (first detection device)
64 pressure sensor (second detector)
65 Controller (Control Device)
65a pump displacement conversion block 65b torque calculation block 65c pressure conversion block 65d current conversion block 102 variable displacement main pump (first pump device)
112 Regulator (1st pump controller)
112b LS control valve (1st load sensing control unit)
112c LS control piston (1st load sensing control unit)
112d, 112e, 112f Torque control piston (first torque control unit)
202 Variable displacement main pump (second pump device)
202a Third Discharge Port 212 Regulator (Second Pump Controller)
212b LS control valve (second load sensing control unit)
212c LS control piston (second load sensing control unit)
212d Torque control piston (second torque control unit)
111, 211, 311 Differential pressure reducing valve 104, 204, 304 Control valve unit 122, 123, 124a, 124b Operating device 131, 132, 133 first, second, third load pressure detection circuit

Claims (2)

And a motor
A variable displacement first pump device driven by the prime mover;
A variable displacement second pump device driven by the prime mover;
A plurality of actuators driven by pressure oil discharged by the first and second pump devices;
A plurality of flow control valves for controlling the flow of pressure oil supplied from the first and second pump devices to the plurality of actuators;
A plurality of pressure compensation valves that respectively control the differential pressure of the plurality of flow control valves;
A first pump control device that controls a discharge flow rate by controlling a capacity of the first pump device;
And a second pump control device that controls the discharge flow rate by controlling the capacity of the second pump device.
In the first pump control device, when at least one of the discharge pressure and the volume of the first pump device increases and the absorption torque of the first pump device increases, the absorption torque of the first pump device is a first maximum. A first torque control unit for controlling the displacement of the first pump device so as not to exceed the torque;
In the second pump control device, when at least one of the discharge pressure and the volume of the second pump device increases and the absorption torque of the second pump device increases, the absorption torque of the second pump device is the second maximum. A second torque control unit for controlling the displacement of the second pump device so as not to exceed the torque, and a discharge pressure of the second pump device when the absorption torque of the second pump device is smaller than the second maximum torque And a load sensing control unit for controlling the capacity of the second pump device so that the target differential pressure is higher than the maximum load pressure of the actuator driven by the pressure oil discharged by the second pump device. In the drive,
A variable valve device which is provided in the second pump device and changes the setting according to the volume of the second pump device to generate a pressure according to the volume of the second pump device;
A first detection device that detects the discharge pressure of the second pump device;
A second detection device for detecting the pressure generated by the variable valve device;
Proportional solenoid valve,
A torque control piston for guiding the output pressure of the proportional solenoid valve and reducing the volume of the first pump device by the increase of the output pressure;
The absorption torque of the second pump device is calculated from the discharge pressure of the second pump device detected by the first detection device and the output pressure of the variable valve device detected by the second detection device, and this second pump device And a controller for outputting a drive signal corresponding to the absorption torque of
The proportional solenoid valve operates according to the drive signal output from the control device, and outputs a pressure corresponding to the absorption torque of the second pump device to the torque control piston.
The torque control piston reduces the first maximum torque by the absorption torque of the second pump device by controlling the capacity of the first pump device by the pressure according to the absorption torque of the second pump device. A hydraulic drive system for a construction machine characterized by In the hydraulic drive system for a construction machine according to claim 1,
The second pump device has a variable displacement member that changes the displacement of the second pump device by changing the tilt angle.
The variable valve device is a variable pressure reducing valve whose output pressure is variable according to the capacity of the second pump device.
Said variable pressure reducing valve and a setting spring biasing the pressure reducing direction, and a piston for changing the urging force of the setting spring, said piston abutting said variable capacitance element of the second pump device, the variable displacement displaced in accordance with the tilt angle of the member to change the biasing force of the set spring, the variable pressure reducing valve reduces the pressure in the pressure of the pilot pressure oil supply path in response to changes in the biasing force of the set spring, the vacuum A hydraulic drive system for a construction machine, characterized in that the pressure is output.

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