JPH0384202A - Hydraulic driving unit for construction machine - Google Patents

Abstract

PURPOSE:To decrease the power loss and improve the operability by providing an indicating means for setting the flow increase speed of pressure oil to be supplied to a first actuator and driving a first flow control valve by the signal of an operating means in such a way as to make its action sped lower than the action speed target value operated according to the indication of the indicating means. CONSTITUTION:A controller 43 receives the input of the command signal Es for setting the flow increase speed of pressure oil to be supplied to a swing motor 21 from an indicating device 39 operated by an operator, the signal of a differential pressure detector 42 for detecting the maximum load pressure of the swing motor 21 and a boom cylinder 22, and the operating signal Esw of a control lever device 29. The controller 43 then operates the action speed target value of a flow control valve 23, outputs the operating signals of solenoid proportional pressure reducing valves 33, 34 and controls in such a way as to make the action speed of the flow control valve 23 less than the target value. The power loss is therefore decreased and the operability can be improved.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a hydraulic drive system for construction machinery such as a hydraulic excavator, and particularly relates to a swing motor that drives a rotating body of a hydraulic excavator, a boom cylinder that drives a boom, etc. The present invention relates to a hydraulic drive system for construction machinery suitable for performing combined operation of an actuator that drives an inertial load and an actuator that drives other loads.

[Conventional technology]

In recent years, in hydraulic drive systems for construction machinery such as hydraulic excavators and hydraulic cranes that are equipped with multiple hydraulic actuators that drive multiple driven objects, the discharge pressure of the hydraulic pump is controlled in conjunction with the load pressure or required flow rate. At the same time, a pressure compensation valve is arranged in conjunction with the flow control valve, and the pressure compensation valve controls the differential pressure across the flow control valve, thereby stably controlling the supply flow rate during combined drive. There is. this house,
Load sensing control is a typical example of controlling the discharge pressure of a hydraulic pump in conjunction with load pressure.

Load sensing control is to control the discharge amount of a hydraulic pump so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of multiple hydraulic actuators by a certain value. Economical operation is possible by increasing or decreasing the discharge amount of the hydraulic pump.

By the way, since the discharge amount of the hydraulic pump has an upper limit, that is, the maximum possible discharge amount, when the hydraulic pump reaches the maximum possible discharge amount during combined driving of a plurality of actuators, a state in which the pump discharge amount is insufficient occurs. This is commonly known as hydraulic pump saturation. When this saturation occurs, the pressure oil discharged from the hydraulic pump flows preferentially to the actuator on the low pressure side, and sufficient pressure oil is no longer supplied to the actuator on the high pressure side, allowing multiple actuators to be driven in a desired manner. I won't be able to do that.

In order to solve these problems, DE-Al-3422
165 (corresponding to JP-A-60-11706), U.S. Patent No. 4,739,617, etc.,
For each pressure compensation valve that controls the differential pressure across the flow control valve, a spring is used to set the target value for the differential pressure across the flow control valve. The control force based on the control force is applied directly or indirectly. With this configuration, when saturation occurs in the hydraulic pump, the differential pressure between the pump discharge pressure and the maximum load pressure decreases, so the target value of the differential pressure across the flow control valve in each pressure compensation valve also decreases. As a result, the pressure compensation valve associated with the low-pressure side actuator is further throttled, and pressure oil from the hydraulic pump is prevented from flowing preferentially to the low-pressure side actuator. Thereby, the pressure oil from the hydraulic pump is divided in accordance with the ratio of the required flow rate (valve opening degree) of the flow rate control valve and is supplied to the plurality of actuators, making it possible to perform appropriate combined driving.

In addition, in this specification, the function of the pressure compensation valve is referred to as the function of the pressure compensation valve that allows the pressure oil from the hydraulic pump to be reliably divided and supplied to a plurality of actuators regardless of the discharge state of the hydraulic pump. For convenience, this is referred to as "divided flow compensation" and the pressure compensation valve is referred to as "divided flow compensation valve."

[Problem to be solved by the invention]

By the way, in this conventional hydraulic drive device, when the plurality of actuators include an actuator that drives an inertial load and an actuator that drives a normal load, when the two actuators are combinedly driven, there is a difference in the load pressure between the two actuators. There were the following problems. Below, we will explain the problem using a hydraulic excavator as an example of a construction machine, a swing motor that drives a rotating structure as an actuator that drives an inertial load, and a boom cylinder that drives a boom as an actuator that drives a normal load. do.

In a hydraulic excavator, when loading earth and sand onto a truck by driving the swing motor and boom cylinder to perform a combined operation of swinging and raising the boom, the load pressure of the swing motor reaches its maximum at the start of this combined operation. The discharge pressure of the hydraulic pump is controlled to be higher than its maximum load pressure by a certain value by load sensing control and control of the diversion compensation valve, and the discharge amount of the hydraulic pump is controlled by the aforementioned diversion compensation valve (pressure compensation valve). This function distributes the flow rate to the swing motor and boom cylinder according to the ratio of the required flow rate of these flow control valves. At this time, since the swing motor has a large load inertia, it cannot instantaneously reach a speed commensurate with the supplied flow rate, and excess pressure oil, excluding the flow rate that actually flows into the swing motor, flows into the tank via the relief valve. It will leak out and cause a huge loss of power.

In addition, since the boom cylinder is an actuator on the low load pressure side, the boom diversion compensation valve is throttled, and the pressure oil flowing into the boom cylinder is at the pressure equal to the pump discharge pressure minus the boom load pressure. This also causes a significant power loss. For example, if the pump discharge pressure is 259 kg/a, the driving pressure required to raise the boom is approximately 100 kg.
/cnf, and the difference of 150 kg/cnf is throttled by the shunt compensation valve and discarded as heat.

Furthermore, as the discharge amount of the hydraulic pump is distributed between the swing motor and the boom cylinder according to the ratio of the required flow rate of the flow control valve, when the hydraulic pump discharge amount has reached its maximum, the pump discharge amount will not be sufficient for swinging. The flow rate of the pressure oil that is allocated as needed and supplied to the boom cylinder is correspondingly reduced. In addition, hydraulic pumps generally have input torque limit control to prevent overloading the prime mover that drives them, so the discharge volume of the hydraulic pump is commensurate with the high discharge pressure. This results in a low flow rate, and this low flow rate is distributed between the swing motor and the boom cylinder by the shunt compensation valve. As a result, the flow rate of the pressure oil supplied to the boom cylinder further decreases, the driving speed of the boom cylinder is regulated, and the combined operation results in a small lifting amount of the boom, which significantly impairs work efficiency.

The above is a problem when an inertial load (swing) and a normal load (boom) are operated in combination, but when the actuator that drives the inertial load is started, a part of the pressure oil supplied to the actuator pushes the relief valve. flows into the tank through
The problem of power loss also occurs when an actuator for driving an inertial load, ie, a swing motor, is operated alone.

An object of the present invention is to provide a hydraulic drive system for construction machinery having an actuator that drives an inertial load, which can reduce power loss when driving a driven body related to this actuator. be.

Another object of the present invention is to provide a hydraulic drive system for construction machinery that has an actuator that drives an inertial load and an actuator that drives a normal load, and to reduce power loss and control operation when performing combined driving of a driven body involving these actuators. An object of the present invention is to provide a hydraulic drive device for construction machinery that can improve performance.

[Means to solve the problem]

According to the first aspect of the present invention, the above-mentioned object includes a hydraulic pump, and at least first and second hydraulic actuators driven by pressure oil supplied from the hydraulic pump, each responsive to an operating signal from an operating means. and the first
and first and second flow control valves that respectively control the flow of pressure oil supplied to the second actuator, and first and second flow control valves that respectively control the differential pressure across the first and second flow control valves. 2, the first actuator is an actuator that drives an inertial load,
In the hydraulic drive system for construction machinery, wherein the second actuator is an actuator that drives a normal load, an instruction means for setting a rate of increase in the flow rate of pressure oil supplied to the first actuator, and a setting by the instruction means. An operation speed target value of the first flow control valve is calculated based on the flow rate increase rate determined, and when the first flow control valve is driven by the operation signal of the operation means, the operation of the first flow control valve is adjusted. This is achieved by providing a control means for controlling the speed to be equal to or less than the target operating speed.

In this hydraulic drive device for construction machinery, preferably, a detection means for detecting the drive of the second actuator is further provided, and the control means outputs an operation signal of the operation means related to the first flow rate control valve. and when the detection means detects the drive of the second actuator, a target operating speed of the first flow rate control valve is calculated, and the operating speed is controlled.

Further, according to the second aspect of the present invention, a first hydraulic pump having a displacement variable means, and at least first and second hydraulic actuators driven by pressure oil supplied from the first hydraulic pump. , first and second flow control valves that respectively control the flow of pressure oil supplied to the first and second actuators, and control differential pressures across the first and second flow control valves, respectively. The discharge pressure of the first hydraulic pump is adjusted to the load of the first and second flow compensating valves in response to a pressure difference between the discharge pressure of the first hydraulic pump and the load pressure of at least the first actuator. a discharge amount control means for controlling the discharge amount of the first hydraulic pump so that the discharge amount is higher than the pressure by a predetermined value, the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives an inertial load. In a hydraulic drive system for construction machinery, which is an actuator that drives a normal load, a second hydraulic pump capable of supplying pressure oil to the second actuator separately from the first hydraulic pump; an opening/closing means for controlling communication of a discharge pipe of a second hydraulic pump; a first detecting means for detecting driving of the first actuator; and a rate of increase in the flow rate of pressure oil supplied to the first actuator. and an instruction means for setting a displacement volume of the first hydraulic pump based on a flow rate increase rate set by the instruction means when driving of the first actuator is detected by the first detection means. calculating an operating speed target value of the variable means, and controlling the displacement variable means so that its operating speed is equal to or less than the operating speed target value when controlling the discharge amount of the first hydraulic pump by the discharge amount controlling means; This is achieved by providing a control means for switching the opening/closing means to a cutoff position.

In this hydraulic drive device for construction machinery, preferably, a second detection means for detecting the drive of the second actuator is further provided, and the control means is configured to detect the first and second actuators by the first and second detection means. When the driving of the second actuator is detected, the operation speed of the displacement variable means is controlled and the opening/closing means is switched to the blocking position.

[Effect]

In the first aspect of the present invention configured as described above, the operating speed target value of the first flow control valve is calculated based on the flow rate increase rate set by the indicating means, and the operating speed of the first flow control valve is adjusted. By controlling the operating speed to be below the target value, the rate of increase in the flow rate of the pressure oil supplied to the actuator is proportional to the operating speed of the flow control valve, and the pressure of the pressure oil supplied to the actuator and driving it is increased. That is, since there is a general relationship in which the driving pressure is regulated by the rate of increase in the flow rate of its supply, the rate of increase in the flow rate of the pressure oil supplied to the first actuator is the operating speed of the first flow control valve. The driving pressure of the first actuator is controlled according to the target value, and the driving pressure of the first actuator is equal to or lower than the value corresponding to the target operating speed value. Therefore, if the set value of the flow rate increase rate in the indicating means is appropriately selected, the driving pressure of the first actuator, that is, the load pressure, can be maintained below the relief pressure of the first actuator, and the power loss due to the relief of pressure oil can be maintained. can be avoided. This also applies to the combined operation of the driven bodies related to the first and second actuators, and it is possible to prevent the pressure oil supplied to the first actuator from being relieved at the start of the combined operation.

On the other hand, during combined operation of the two-wave drive body, the load pressure of the first actuator on the high pressure side is maintained below the relief pressure, so the discharge pressure of the hydraulic pump controlled by load sensing is also lower than before. As a result, the pressure drop across the second flow compensating valve is reduced, and the power loss there is also reduced.

Furthermore, by controlling the operating speed of the first flow control valve, the flow rate of the pressure oil supplied to the first actuator is restricted, so when the discharge amount of the hydraulic pump reaches the maximum, the flow rate of the pressure oil supplied to the first actuator is restricted. As the amount of pressure oil supplied to the actuator decreases, the amount of pressure oil supplied to the second actuator increases.

In addition, when input torque limit control is performed to control the discharge amount of a hydraulic pump, the discharge pressure of the hydraulic pump is kept lower than before, so the pump discharge amount becomes a relatively large flow rate commensurate with the low discharge pressure. , since this large flow rate is distributed to the first and second actuators, the second
The flow rate of pressurized oil supplied to the actuator is further increased. By increasing the flow rate of the pressure oil supplied to the second actuator in this way, the drive speed of the second actuator increases, and the amount of movement of the second actuator relative to the first actuator increases, improving operability. will improve.

In the second aspect of the present invention, the target operating speed of the displacement variable means of the first hydraulic pump is calculated based on the flow rate increase rate set by the instruction means, and the operation speed target value of the displacement variable means of the first hydraulic pump is calculated. By controlling the operating speed to be equal to or less than the operating speed target value and switching the opening/closing means to the cutoff position, the operating speed of the displacement variable means of the first hydraulic pump corresponds to the increasing speed of the pump discharge amount, Since the rate of increase in the flow rate of the pressure oil supplied to the first actuator matches the rate of increase in the discharge amount of the first hydraulic pump, the rate of increase in the supply flow rate is determined by the displacement variable means of the first hydraulic pump. The driving pressure of the first actuator is controlled in accordance with the operating speed target value, and the driving pressure of the first actuator is equal to or less than the value corresponding to the calculated operating speed target value. Therefore, in this second aspect of the invention as well, as in the first aspect of the invention, the drive pressure of the first actuator can be maintained below the relief pressure, and power loss can be reduced.

In addition, during combined operation, communication between the discharge pipes of the first and second hydraulic pumps is cut off, so the second hydraulic pump becomes a hydraulic pump exclusively for the second actuator.
The pressure drop at the second flow compensating valve is reduced, and power loss can be further reduced. Further, since a sufficient flow rate of pressure oil can be supplied to the second actuator from the second hydraulic pump, the driving speed of the second actuator is increased and operability is improved.

〔Example〕

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

First Embodiment First, a first embodiment of the present invention will be described with reference to FIGS. 1 to 6.

In FIG. 1, the hydraulic drive system of this embodiment is shown as being applied to a hydraulic excavator, and includes one variable displacement hydraulic pump, that is, a main pump 20 driven by a prime mover (not shown), and a main pump 20. A plurality of actuators driven by pressure oil discharged from 20, that is, a swing motor 21 and a boom cylinder 22, and a flow control valve that controls the flow of pressure oil supplied to each of these plurality of actuators, that is, a swing direction. The switching valve 23 and the boom directional switching valve 24 are arranged upstream of these flow control valves to control the differential pressure generated between the inlet and outlet of the flow control valve, that is, the differential pressure across the flow control valve. It is provided with pressure compensation valves to control, ie branch compensation valves 25,26.

In addition, a relief valve 28 is provided in the discharge pipe line 27 of the main pump 20.
is connected, and the relief valve 28 causes the pressure oil from the main pump 20 to flow out into the tank 30 when it reaches a set value, thereby preventing the pump discharge pressure from exceeding the set value.

The discharge amount of the main pump 20 is determined by the discharge amount control device 31 so that the discharge pressure Ps is the load pressure on the high pressure side of the swing actuator 21 and the boom cylinder 22 (hereinafter referred to as the maximum load pressure) pHl11! Load sensing control is performed so that the load becomes higher by a predetermined value ΔP LSO. The discharge amount control device 31 is driven by a tilt drive device 31a that drives the displacement variable means of the main pump 20, that is, the swash plate 20a, and increases or decreases the displacement, and an electric control signal, and is driven by an electric control signal. The control pressure is comprised of an electromagnetic proportional pressure reducing valve 31b that generates control pressure based on the pilot pressure and outputs it to the tilt drive device 31a to adjust its displacement.

The flow control valves 23 and 24 are pilot-operated valves driven by pilot pressure, and the flow control valves 23 and 24 are pilot-operated valves driven by pilot pressure.
3, an operating lever device 29 is provided which generates an electrical command signal ESW according to the operation amount, and is driven by the electrical operation command signals ESR and ESL, and generates a pilot pressure according to this. Solenoid proportional pressure reducing valve 33.3
4 are provided. The flow rate control valve 24 is provided with a control lever device (not shown) including a pilot valve that generates a pilot pressure according to the amount of operation.

The branch compensating valves 25 and 26 each have a configuration in which the outlet pressure and inlet pressure of the flow control valves 23 and 24 are guided, and a control force based on the differential pressure before and after is applied in the valve closing direction, as in a normal pressure compensating valve. and a spring 25 that acts in the valve opening direction.
a, 26a, and drive parts 25b, 26b to which the control pressure output from the electromagnetic proportional pressure reducing valve 35, 36 is guided and which applies a corresponding control force in the valve closing direction. By changing the control pressure output from .36, the set values of the springs 25a and 25b are changed, and the target value of the differential pressure across the flow control valves 23 and 24 is changed.

A relief valve 37 for rotation is provided in the drive circuit of the rotation motor 21.
A shuttle valve 38 for detecting the maximum load pressure of the swing motor 21 and the boom cylinder 22 is connected to the flow control valves 23 and 24, respectively.

The hydraulic drive device of this embodiment also includes an instruction device 39 that is operated by an operator from the outside and outputs an electrical command signal EI for setting the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21; It is equipped with a differential pressure detector 42 that introduces the discharge pressure P$ of the main pump 20 and the maximum load pressure PIIIK of the swing motor 21 and boom cylinder 22, and detects the differential pressure ΔPLS between the two.

Then, the operation signal E SW of the operation lever device 29.

The command signal Es of the indicating device 39 and the detection signal of the differential pressure detector 42 are input to the controller 43, and based on these signals, an operation command signal E for the electromagnetic proportional pressure reducing valves 33 and 34 is sent.
SR, ESL and electromagnetic proportional pressure reducing valve 31b, 35.36
A control command signal for is calculated and generated.

In this embodiment, as shown in FIG. 2, the indicating device 39 consists of a voltage setting device including a variable resistor 44, and when the operator changes the position of the movable contact, the voltage is set at a corresponding level. Ru. This voltage value is taken into the controller 43 as a command signal Ex, and the command signal Es is A/D converted in the controller 43 and then sent to the CPU. In the CPU, the increment Δ per cycle is calculated according to the procedure shown in the flowchart in FIG.
Calculate E. That is, in step S1, the command signal E
Read the A/D conversion value of $, and in step S2, Δ
Set E=A/D conversion value and find the increment ΔE per cycle. In this embodiment, this increment ΔE is equal to the flow rate control valve 23.
The flowchart in Figure 3 corresponds to the operating speed target value of
After all, the procedure for calculating the operating speed target value of the flow rate control valve 23 is shown. This increment ΔE is determined by the operation command signal E for the electromagnetic proportional pressure reducing valve 33, 34 in the controller 43.
It is used to find SR and ESL.

The contents of the calculations performed by the controller 43 are shown in a flowchart in FIG. This flowchart shows the electromagnetic proportional pressure reducing valve 3.
3.34 shows the calculation procedure of the operation command signals ESR and ESL.

First, in step SIO, the operation signal Esv and command signal E3 are read. Next, in step Sll, it is determined whether the operation position of the flow rate control valve 23 indicated by the operation signal Esv is the neutral position, the left direction, or the right direction, and if it is the neutral position, the process proceeds to step 812. ,
If the direction is to the left, the process proceeds to steps S1 and 3; if the direction is to the right, the process proceeds to step SI4.

In step 812, since the operation signal ESW indicates the neutral position, ESR=0 for each of the operation command signals ESR and ESL. Set ESL=0.

In step S13, since the operation signal ESW instructs the left direction, the operation command signal ESR is
In addition to setting R=0, the operation command signal E which is stored in advance
From the functional relationship between SW and the operation command target value E SLO of the stroke amount of the flow rate control valve 23, E5LO=fSW(ES
W) and find the operation command target value E SLO corresponding to the operation signal ESW at that time. This ESLO is shown in Figure 5.
The functional relationship of = f SW (ESW) is shown.

Then, the process advances to step 815 and ESLO>ESL-1
Determine whether +ΔE. Here, ESL-1 is the operation command signal obtained in the previous control cycle, and ΔE is the increment per cycle based on the command signal Es mentioned above. It is determined whether the added value exceeds the operation command target value ESLO. If it is determined that E5C-1+ΔE does not exceed E SLO, the process proceeds to step 316, and E S
Set L = E 5L-1 + ΔE, and add the increment ΔE to the previous operation command signal E 5L-1 to make the operation command signal E.
SL, and if it is determined that E5L-1+ΔE exceeds ESt,0, the process advances to step S17, and ESL=ES
LO is set as the operation target value E. SLO is the operation command signal ES.
It is considered to be L.

On the other hand, in step 814, since the operation signal ESW instructs the right direction, the operation command signal ESL is set to ESL=0, and the operation command signal ESW and the flow rate which are stored in advance are set as 0 as in step S13. control valve 2
Based on the functional relationship between the stroke amount and the operation command target value ESRO in step 3, let ESRO=fSW (ESW), and find the operation command target value ESRO corresponding to the operation signal ESW at that time. The functional relationship of ERLO=f SW (ESW) is shown in FIG.
) is the same as the functional relationship. And below step 31
The process proceeds to steps 8 to 820, where calculations similar to steps 315 to SI7 are performed to obtain the operation command signal ESL.

After determining the operation command signals ESR and ESL as described above, the process advances to step S21, where E5R-1=ESR, E5L-1= is calculated for the next control cycle.
E_SL, and in step 322, the operation command signal E_SR.

Output ESC to electromagnetic proportional pressure reducing valves 33 and 34.

The controller 43 further obtains a control command signal for the electromagnetic proportional pressure reducing valve 31b of the discharge amount control device 31 and a control command signal for the electromagnetic proportional pressure reducing valve 35, 36 of the branch compensation valve 25, 26. Since this is not a main configuration in this embodiment, its calculation procedure is not shown in the flowchart, but roughly speaking, for the electromagnetic proportional pressure reducing valve 31b, the main pump 2 discharge pressure Ps and maximum load pressure Pllll5 of the swing motor 21 and boom cylinder 22! The differential pressure ΔPLS with the predetermined value ΔPL
A discharge amount target value for maintaining SO is calculated, and a signal corresponding to this discharge amount target value is outputted to the electromagnetic proportional pressure reducing valve 31b as a control command signal.

In addition, for the electromagnetic proportional pressure reducing valves 35 and 36, based on the detection signal of the differential pressure detector 42, when the discharge amount of the main pump 20 is saturated and the differential pressure ΔPLS becomes less than a predetermined value ΔP LSO, the shunt compensation valve is activated. 25.26 is calculated and outputted as a control command signal to the electromagnetic proportional pressure reducing valve 35.36.

Next, the operation of this embodiment configured as above will be explained.

First, when no flow control valve is operated and the actuator is not driven, the operation signal ESW of the operation lever device 29 is not input to the controller 43, so step St in the flowchart of FIG.
At t, it is determined that the operation signal ESW is neutral, and at step 812, ESR=0, ESL=
It is set to 0. Therefore, the flow rate control valve 23 is held at the neutral position. The flow rate control valve 24 is also maintained at the neutral position, with no pilot pressure being generated, since the corresponding operating lever device (not shown) is at the neutral position.

Next, consider the independent turning operation. In this case, prior to operating the operating lever device 29, the operator first operates the indicating device 39 to set the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21. That is, by operating the indicating device 39, an electrical command signal E@ of a desired level is output. As described above, the controller 43 receives this command signal Es and calculates the increment ΔE1 per cycle corresponding to the command signal Es, that is, the operating speed target value of the flow rate control valve 23.

Next, operate the operating lever device 29 to output the operating signal ESW.
Output. In the controller 43, the direction indicated by the operation signal ESW is determined in step Sll of the flowchart of FIG. Here, if the indicated direction is the left direction, then in step S13 E SR = 0
is calculated, and in steps 813 to 817 E SL=
E SLO or E SL = E5L-1+ΔE is determined. Then, these operation command signals ESR and ESL are output to the electromagnetic proportional pressure reducing valves 33 and 34, and the flow control valve 23 gradually begins to open at a speed corresponding to the increment ΔE. is supplied to the swing motor 21 at a flow rate increase rate corresponding to the increment ΔE. As a result, the swing motor 21 is driven with an acceleration corresponding to the increment ΔE.

Here, the relationship between the time t during the turning operation, the operation command signal ESR, and the command signal Es (increment ΔE) is shown in FIG.

After the start of the turn, the operation command signal ESR increases at a gradient corresponding to the command signal Es. The slope becomes larger as the command signal Es, that is, the increment ΔE becomes larger. This gradient also corresponds to the operating speed of the flow control valve 23, and further corresponds to the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21.
That is, it corresponds to the driving acceleration of the swing motor 21. Then, the operation command signal ESR increases until it finally matches the operation command target value E SLO determined from the operation signal ESW.

Next, let us consider a combined operation of turning and boom, such as a combined operation of turning and raising boom, which is performed when loading earth and sand onto a truck. In this case, the flow rate control valve 23 for swinging is controlled in the same way as the independent swinging operation described above, and the swinging motor 21 is controlled in the same way. For the boom cylinder 22, a flow control valve 24 is driven by an operation lever device (not shown), and pressure oil is supplied to the boom cylinder 22 via the flow control valve 24, thereby raising the boom. At this time, if the discharge amount of the main pump 20 reaches the maximum, a control command signal is output from the controller 43 to the electromagnetic proportional pressure reducing valve 36, and the drive section 26b of the branch compensation valve 26
Applying control pressure to forcefully throttle the branch compensating valve 26,
The flow rate ratio is maintained in accordance with the opening ratio of the flow rate control valves 23 and 24.

As described above, in this embodiment, the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21 can be arbitrarily set by operating the indicating device 39. The pressure of the pressure oil that drives this, that is, the driving pressure, can be lower than the relief pressure of the relief valve 37 for swinging, suppressing the outflow of pressure oil from the relief valve 37, and reducing power loss due to acceleration of the swing. can be reduced.

Furthermore, during combined operation of swing and boom, the drive pressure of the swing motor 21, which is on the high-load pressure side, is lower than the swing relief pressure, so that the discharge pressure of the main pump 20, which is controlled by load sensing, is also lower than before. For this reason,
The pressure drop at the boom-side shunt compensation valve 26 is also reduced,
Furthermore, power loss can be reduced.

In addition, during combined operation of swing and boom, the flow control valve 2
By controlling the operating speed of No. 3 and limiting the flow rate of pressure oil supplied to the swing motor 21, when the discharge amount of the main pump 20 reaches the maximum, the amount of pressure oil supplied to the swing motor 21 can be controlled. In accordance with the decrease, the supply flow rate of pressure oil distributed to the boom cylinder 22 increases. In addition, when input torque limitation control is performed when controlling the discharge amount of the main pump 20,
Since the discharge pressure of the main pump 20 is maintained lower than before, the pump discharge amount becomes a relatively large flow rate commensurate with the low discharge pressure, and this large flow rate is distributed, so that the pressure oil supplied to the boom cylinder 22 The flow rate of will further increase. By increasing the flow rate of pressurized oil supplied to the boom cylinder 22 in this way, the driving speed of the boom cylinder 22 increases, and the boom can be lifted by a large amount in a combined operation of turning and raising the boom. can significantly improve sex.

Modification of the First Embodiment A first modification of the first embodiment will be explained with reference to FIGS. 7 and 8. This embodiment shows a modification of the pointing device.

In FIG. 7, the indicating device 39A has four contacts A-D.
It consists of a switching device including a movable contact 50 for. Contacts A-C are connected to the CPU in the controller 43A.
are connected to the input terminals Di 1 + D + 2 + Di 3 of the input terminals Di 1 . Di 2゜Di
3 is connected to a power supply via resistors 51a, 51b, and 51c. With such a configuration, when the movable contactor 50 is in a position where it contacts the contact point C as shown in the figure, for example, the input terminal Di 1 is grounded and the voltage becomes O, and the other input terminals Di 2 . D3 is maintained in a state where the power supply voltage is applied.

In the controller 43A, input terminals Di1 and Di
2. The flow rate increase rate, that is, the increment ΔE is set as shown in FIG. 8 according to the voltage state of Di 3. Step 8
In step S30, it is determined whether the voltage at the input terminal Di3 is O or not. If it is 0, in step S31, the electromagnetic proportional pressure reducing valve 33. 34 (see FIG. 1), the increment ΔE per cycle of the operation command signals ESR and ESL is set to a pre-stored value ΔEA. If the voltage at the input terminal Di3 is not O, the process advances to step 832 and the voltage at the input terminal Di3 is not O.
It is determined whether the voltage of 2 is O, and if it is 0, the increment ΔE is set to a pre-stored value ΔEB in step 333.

If the voltage at input terminal Di2 is not 0, step S
The process proceeds to step S34, and it is determined whether the voltage at the input terminal Di1 is O. If it is 0, the increment ΔE is set to a pre-stored value ΔEC in step S35. Finally, input terminal Di
If the voltage of 1 is not O, proceed to step 836;
Set the increment ΔE to a pre-stored value ΔED.

By switching the position of the movable contactor 50 as described above, the increment ΔE can be set according to the position.

Next, a second modification of the first embodiment will be explained with reference to FIG. 9. In FIG. 9, the same steps as those in the processing procedure shown in FIG. 4 are given the same reference numerals. In this embodiment, the flow rate increase speed control for the swing motor 21 is performed only during a combined operation of swing and boom raising.

In the hydraulic drive system of this embodiment, as shown by imaginary lines in FIG.
A drive detector 53 is further provided, which is connected to the pilot line 52 connected to the drive section on the side corresponding to boom raising, and detects that pilot pressure is applied to the drive section, and the detection signal EB is sent to the controller 43. is output to.

In the controller 43, step S shown in FIG.
In addition to the operation signal ESW and the command signal ES, the IOA further reads the drive detection signal EB from the drive detector 53, and determines whether or not the drive detection signal EB has been input in step 823 or 324. Proceeds to step S15 or S18 only when the ES
LO>ESL-1+ΔE or E SRO>E 5R-
When it is determined that 1+ΔE, step 816. 819
The operation command signal ESL or ESR is obtained by adding the increment ΔE to the previous operation command signal E5L-1 or ESR-1.

According to this embodiment, the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21 is controlled only during the combined operation of swing and boom raising, and the same effect as the first embodiment can be obtained, and the swing When operated independently, the swing motor 21 can be driven as before.

Second Embodiment A second embodiment of the present invention will be explained with reference to FIGS. 10 to 13.

FIG. 10 is a diagram showing an all-opposed hydraulic drive system of this embodiment, and in the figure, the same elements as those shown in FIG. 1 are given the same reference numerals.

In FIG. 10, the hydraulic drive system of this embodiment includes, in addition to the first hydraulic pump 20, a variable displacement second hydraulic pump,
That is, the main pump 60 is provided, and the discharge pipe 6 of the main pump 60 is
1 is also provided with a relief valve 62. Main pump 60
Similar to the main pump 20, the discharge amount is controlled by a discharge amount control device 63 consisting of a tilt drive device 63a that drives the swash plate 60a and an electromagnetic proportional pressure reducing valve 63a that is driven by the control command signal Ep2. is the turning actuator 21
and the load pressure on the high pressure side of the boom cylinder 22 (hereinafter referred to as the maximum load pressure) Pu+sx to a predetermined value ΔP L
Load sensing control is performed so that SO becomes higher.

The electromagnetic proportional pressure reducing valve 31b of the discharge amount control device 31 for the main pump 20 is driven by the control command signal Epl.

An electromagnetic on-off valve 65 is connected to a pipe line 64 that connects the discharge pipe line 27 of the main pump 20 and the discharge pipe line 61 of the main pump 60.
However, when the drive signal Ed is applied, the position switches to the position where communication with the pipe line 64 is cut off, and the
Separate the discharge oil of the two main pumps 20.60.

The swing flow control valve 23 is a pilot-operated valve like the first embodiment, but unlike the first embodiment, it is driven by an operating lever (not shown) like the boom flow control valve 24. This is done by means of pilot pressure generated by the pilot valve of the device.

As shown in the figure, drive detectors 65 and 66 are connected to the pilot line, which detect the application of pilot pressure to the pilot line and output drive detection signals ESIIR and ESIIL. Flow control valve 2
A swing load pressure switching valve 68 is provided in the load line between 3 and the shuttle valve 38, and normally sends the swing load pressure to the shuttle valve 38, but when the drive signal Ed is applied, it switches and changes the tank pressure. to the shuttle valve 38.

In the differential pressure detector 42, the pump discharge pressure total introduction ta pipe is connected to the main pump 60 pipe line through the electromagnetic on-off valve 65.
When the electromagnetic on-off valve 65 is closed, the discharge pressure of the main pump 60 is introduced. A second differential pressure detector 69 is connected to the discharge line 27 of the main pump 20 and the load line of the swirl flow control valve 23. As a result, the electromagnetic on-off valve 65
When the swing load pressure switching valve 68 is switched from the illustrated position by the drive signal Ed, the differential pressure between the discharge pressure of the main pump 20 and the load pressure of the swing motor 21 is detected by the differential pressure detector 69, and the differential pressure The detector 42 detects the differential pressure between the discharge pressure of the main pump 60 and the load pressure of the boom cylinder 22 . The detection signal of differential pressure detector 60 is represented by E dpl, and the detection signal of differential pressure detector 42 is represented by E dp2.

Further, as in the first embodiment, an instruction device 39 is operated from the outside by an operator and outputs an electrical command signal E$ for setting the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21. It is provided.

Then, the command signal Es of the indicating device 39, the drive detector 66
．． 67 detection signals ESdR, ESdL, differential pressure detector 6
The detection signals Edpl and Edp2 of 0.42 are input to the controller 70, and based on these signals, the control command signals Epl and Epl for the electromagnetic proportional pressure reducing valves 31b and 63b are
Control command signals for Ep2 and the electromagnetic proportional pressure reducing valves 35 and 36 are calculated and generated.

The configuration of the indicating device 39 and the procedure for calculating the increment ΔE from the command signal E$ are the same as in the first embodiment, but in this embodiment, the increment ΔE is the tilting speed target value of the swash plate 20a of the main pump 20. Corresponding to this, the controller 70 sends a control command signal E to the electromagnetic proportional pressure reducing valve 31b of the discharge amount control device 31.
° used to determine pl.

The contents of the calculations performed by the controller 70 are shown in a flowchart in FIG. This flowchart shows the calculation procedure of control command signals Epl and Ep2 for the electromagnetic proportional pressure reducing valves 31b and 63b of the discharge amount control device 31.63.

First, in step 34G, the command signal Ex, drive detection signals ESdR, ESdL, and differential pressure detection signal Edpl are output.

Load E dp2. Next, in step S41, the drive detectors 66, 67 (7) detection signals ESdR,E
It is determined whether or not SdL is present, that is, whether or not driving of the swing motor 21 is commanded, and the detection signal ESdR is determined.
, ESdL, that is, when the driving of the swing motor 21 is not instructed, the process advances to step S42. In step S42, a control command signal Epl for load sensing control of the main pump 20.60 based on the detection signal Edp2 from the differential pressure detector 42;
E92 is calculated. This method of determination is as follows: - As an example, the first
Load sensing differential pressure ΔPLS and control command signal Epl as shown in Figure 2.

A functional relationship with Ep2 is stored in advance, and a control command signal corresponding to the differential pressure ΔPI, S indicated by the detection signal Edp2 is determined from this functional relationship. In FIG. 12, ΔPLSO is the target differential pressure. Next, the process proceeds to step 843, where Epl-1=Epl is set for calculation in the next control cycle.

If it is determined in step S41 that there are detection signals ESdR and ESdL, that is, if driving of the swing motor 21 is commanded, the process proceeds to step S44, and the electromagnetic on-off valve 65
and outputs a drive signal Ed to the swing load pressure switching valve 68,
Each of these valves is switched from the position shown.

Next, the process proceeds to step S45, where a control command signal Ep2 for load sensing control of the main pump 60 is calculated based on the detection signal Edp2 from the differential pressure detector 42. Similar to step 342, this determination is performed using the functional relationship between the differential pressure ΔPLS and the control command signal E dp2 shown in FIG. Furthermore, the process proceeds to step 846, where a control target value EplG for load sensing control of the main pump 20 is calculated based on the detection signal Edpl from the second differential pressure detector 69. This calculation method is the same as step 842, and the differential pressure ΔPLS and control target value Ep are similar to those shown in FIG.
From the functional relationship with lO, the differential pressure ΔPL of the detection signal E dpl
A control target value EplO corresponding to S is determined.

Next, the process advances to step S47, and EplO>Epil+
It is determined whether ΔE. Here, Epl-1 is the control command signal obtained in the previous control cycle, and ΔE is the increment per cycle calculated based on the above-mentioned command signal Ex. Therefore, it is determined here whether the value obtained by adding the increment ΔE to the previous control command signal Epl-1 exceeds the control command target value EplO. Here, if it is determined that Epl-1+ΔE does not exceed EplO, the process proceeds to step 848, and Ept=Epit+Δ
E, and add an increment ΔE to the previous control command signal E pt-t.
The value obtained by adding up is set as the control command signal Epl, and Epl-1+
If it is determined that ΔE exceeds EpHl, step 849
, set Epl=Ep10, and set the control target value E SL
O is taken as the control command signal Epl. and step S5
0, and similarly to step S43, set Epi-1=E pi for calculation in the next control cycle.

As described above, the control command signals E pl, E p2
After determining the control command signal Ep, the process proceeds to step 851 and the control command signal Ep
l.

Ep2 is the electromagnetic proportional pressure reducing valve 3 of the discharge amount control device 31.63
1b and 63b.

Next, the operation of this embodiment configured as above will be explained.

First, consider the independent operation of the boom 22. In this case, when the boom control lever device (not shown) is operated, the boom flow control valve 24 opens, and the pressure oil from the two main pumps 20 and 60 passes through the flow control valve 24 to the boom cylinder 22.
supplied to At this time, since there are no detection signals ESdR and ESdL from the drive detectors 66 and 67, the controller 70 performs step 8 in the flowchart of FIG.
41, the process proceeds to step 842, where control command signals Epl and Ep2 for load sensing control of the main pump 20.60 are calculated based on the detection signal Edp2 from the differential pressure detector 42, and the discharge It is output to the electromagnetic proportional pressure reducing valves 31b and 63b of the quantity control device 31.63. As a result, the discharge amount of the main pump 20°60 is
The discharge pressure is controlled to be higher than the load pressure of the boom cylinder 22 by a predetermined value Δp tso.

Next, consider the independent turning operation. In this case, prior to operating the swing operation lever device, the operator first operates the indicating device 39 to set the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21.

That is, by operating the indicating device 39, an electrical command signal Es of a desired level is output. The controller 43 receives this command signal E$ and calculates the increment ΔE1 per cycle corresponding to the command signal Es, that is, the target value of the tilting speed of the swash plate 20a of the main pump 20, as described above.

Next, the operating lever device for swing is operated to open the flow control valve 23 to a desired opening degree, and pressure oil is supplied to the swing motor 21 through the flow control valve 23. At this time, the drive detector 6
Since there is either the detection signal ESdR or ESdL from 6.67, the controller 70 makes a YES determination in step 841 of the flowchart in FIG. It is output to the load pressure switching valve 68, the discharge pipes 27 and 61 are separated, and the load line of the flow rate control valve 23 and the shuttle valve 38 are separated.

Next, in step 845, a control command signal Ep2 for load sensing control of the main pump 60 is calculated based on the detection signal Edp2 from the differential pressure detector 42,
In step 846, a control target value Epie for load sensing control of the main pump 20 is calculated based on the detection signal Edpl from the second differential pressure detector 69.

Furthermore, the process advances to steps 847 to S49, and E pl=E
plo or E pl = E pl- 10 ΔE is determined. Then, the control command signal E obtained in step 845
p2 and the control command signal Epl obtained in step 848 or 849 is the electromagnetic proportional pressure reducing valve 3 of the discharge amount control device 31.63.
1b and 63b. As a result, the main pump 20
The swash plate 20a gradually increases its tilt angle at a speed corresponding to the increment ΔE, and accordingly, the discharge amount of the main pump 20 increases at a speed corresponding to the increment ΔE. Pressure oil is supplied at a corresponding rate of increase in flow rate.

Here, the relationship between the time t during the turning operation, the control command signal El11, and the command signal Es (increment ΔE) is shown in FIG. After the start of the turn, the control command signal Epl increases at a gradient corresponding to the command signal Ez. The slope is the command signal Es,
That is, it increases as the increment ΔE increases.

This gradient also corresponds to the tilting speed of the swash plate 20a of the main pump 20, and further corresponds to the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21, that is, the driving acceleration of the swing motor 21.

Then, the control command signal Epl is finally the differential pressure signal E
It increases until it matches the control target value E plO determined from dpl.

On the other hand, in the main pump 60, since the boom flow control valve 24 is not opened, the load pressure of the boom cylinder 22 is zero, and the main pump 60 discharge pressure is equal to the target differential pressure Δ.
The discharge amount is controlled to be the minimum so as to almost match the P LSO.

Next, let us consider a combined operation of turning and boom, such as a combined operation of turning and raising boom, which is performed when loading earth and sand onto a truck. In this case, the main pump 20 is controlled in the same way as the above-described independent swing operation, and the swing motor 21 is controlled in the same way. The flow control valve 2 is connected to the boom cylinder 22 by an operation lever device (not shown).
4 is opened, pressure oil is supplied to the boom cylinder 22 via this flow control valve 24, and the boom is raised.

At this time, as described above, in step S45 of the flowchart in FIG. 11, the control command signal Ep2 for load sensing control of the main pump 60 is calculated based on the detection signal Edp2 from the differential pressure detector 42, and Since the pressure is output to the pressure reducing valve 63b, the discharge amount of the main pump 60 is controlled so that the discharge pressure is higher than the load pressure of the boom cylinder 22 by a predetermined value ΔPLSO, and the flow rate control valve 24 is output to the boom cylinder 22. A flow rate is supplied according to the opening degree.

As described above, in this embodiment, the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21 can be arbitrarily set by operating the indicating device 39. The pressure of the pressure oil that drives this, that is, the driving pressure, can be lower than the relief pressure of the relief valve 37 for swinging, suppressing the outflow of pressure oil from the relief valve 37, and reducing power loss due to acceleration of the swing. can be reduced.

In addition, when performing combined swing and boom operations, the oil discharged from the two main pumps 20.60 is separated, and the boom cylinder 20.
2 is supplied with pressure oil from a dedicated main pump 60, so the discharge pressure of the main pump 60 can be lowered without being affected by the load pressure of the swing motor 21, and Pressure drop is reduced and power loss can be further reduced.

In addition, during combined operation of swing and boom, pressure oil is supplied to the boom cylinder 22 from a dedicated main pump 60, so a sufficient flow rate can be obtained, and the boom cylinder 22
Since the driving speed of the boom is increased and the boom can be lifted by a large amount in a combined operation of turning and raising the boom, operability can be significantly improved.

Modification of the second embodiment Next, a modification of the second embodiment will be explained with reference to FIG. In FIG. 14, the same steps as those in the processing procedure shown in FIG. 11 are given the same reference numerals. This embodiment corresponds to the modification shown in FIG. 9 of the first embodiment, in which the flow rate increase speed control for the swing motor 21 is performed only during the combined operation of swing and boom raising. .

In the hydraulic drive system of this embodiment, as shown by the imaginary line in FIG. 10, the side corresponding to the boom raising of the driving section facing from the pilot valve of the boom operation lever device (not shown) to the flow rate control valve 24 A drive detector 53 is further provided, which is connected to a pilot line 52 connected to the drive section to detect that pilot pressure is applied to the drive section, and outputs a detection signal EB to the controller 70.

In the controller 70, in step 840^ shown in FIG.
In addition to R, ESdL, and differential pressure detection signals Edpl and Edp2, the drive detection signal EB from the drive detector 53 is further read, and it is determined in step 852 whether or not the drive detection signal EB has been input, and this is also satisfied. When the control command signal E
Find pi, E p2.

According to this embodiment, the rate of increase in the flow rate of the pressure oil supplied to the swing motor 21 is controlled only during the combined operation of swing and boom raising, and the same effect as the second embodiment can be obtained, and the swing When operated independently, the swing motor 21 can be driven as before.

〔Effect of the invention〕

According to the present invention, the driving pressure of the first actuator that drives the inertial load can be made equal to or lower than the relief pressure.
The pressure oil supplied to the actuator can be suppressed from flowing out from the relief valve, reducing power loss. Further, in the combined drive of the first actuator and the second actuator, the pressure drop at the branch compensating valve related to the second actuator can be reduced, and the power loss in this part can also be reduced.

Furthermore, since the flow rate of the pressure oil supplied to the second actuator can be increased, the amount of movement of the second actuator relative to the first actuator can be increased, and operability can be greatly improved.

[Brief explanation of drawings]

FIG. 1 is a schematic diagram showing the overall configuration of a hydraulic drive system for construction machinery according to a first embodiment of the present invention, FIG. 2 is a schematic diagram showing the configuration of an indicating device, and FIG. 3 is a schematic diagram showing the configuration of an indicating device. FIG. 4 is a flowchart showing the calculation procedure for calculating the increment ΔE based on the command signal E3 of the operation command signal ESR, ES.
This is a flowchart showing the calculation procedure for calculating L, and the fifth
The figure is a diagram showing the functional relationship between the operation signal ESW and the operation command target value E SLO, and FIG. 6 is a diagram showing the relationship between the time t at the start of turning and the operation command signals ESR and ESL.
FIG. 7 is a schematic diagram showing another configuration of the indicating device;
The figure is a flowchart showing the calculation procedure for calculating the increment ΔE from the command signal E$ of the indicating device shown in FIG. 7, and FIG.
10 is a flowchart showing the calculation procedure for calculating
The figure is a schematic diagram showing the overall configuration of a hydraulic drive system for construction machinery according to a second embodiment of the present invention, and FIG. 11 is a flowchart showing the calculation procedure for obtaining control command signals Epl and Ep2 according to this embodiment. , Figure 12 shows the load sensing differential pressure ΔPLS and control command signals Epl and Ep.
2 or control target value EplG, and FIG. 13 shows the relationship between time t at the start of turning and control command signal EplG.
FIG. 14 is a flowchart showing a calculation procedure for obtaining control command signals Epl and Ep2 according to a modification of the second embodiment. Explanation of symbols 20... Main pump (first hydraulic pump) 21... Swing motor (first actuator) 22... Boom cylinder (second actuator) 23... (first)
Flow rate control valve 24... (second) flow control valve 25... (first) branch flow compensation valve 26... (second) branch flow compensation valve 29... Operation lever device 31... Discharge amount control device 39... Indication device 43... Controller (control means) O... Main pump (second hydraulic pump) 5... Electromagnetic on/off valve 6.67... Drive detector 0.・Controller (control means)

Claims (4)

[Claims] (1) a hydraulic pump; at least first and second hydraulic actuators driven by pressure oil supplied from the hydraulic pump; first and second flow control valves that respectively control the flow of pressure oil supplied to the actuator, and first and second branch flow control valves that respectively control the differential pressure across the first and second flow control valves. a compensation valve, wherein the first actuator is an actuator that drives an inertial load, and the second actuator is an actuator that drives a normal load, wherein the first actuator is an actuator that drives a normal load. an instruction means for setting a rate of increase in the flow rate of the supplied pressure oil; and an operation speed target value of the first flow control valve is calculated based on the rate of increase in flow rate set by the instruction means, and an operation speed of the first flow control valve is operated. A hydraulic drive device for construction machinery, comprising a control means for controlling the first flow rate control valve so that its operating speed is equal to or less than the target operating speed when the first flow rate control valve is driven by a signal. . (2) The hydraulic drive device for construction machinery according to claim 1, further comprising a detection means for detecting the drive of the second actuator, and the control means controls the operation means related to the first flow rate control valve. When an operation signal is output and drive of the second actuator is detected by the detection means, an operating speed target value of the first flow control valve is calculated and the operating speed is controlled. Hydraulic drive system for construction machinery. (3) a first hydraulic pump having a displacement variable means; at least first and second hydraulic actuators driven by pressure oil supplied from the first hydraulic pump; First and second flow control valves that respectively control the flow of pressure oil supplied to the actuator, and first and second flow branch compensation that respectively control differential pressures across the first and second flow control valves. a valve; and the first
In response to the differential pressure between the discharge pressure of the hydraulic pump and at least the load pressure of the first actuator, the first hydraulic pressure is adjusted such that the discharge pressure of the first hydraulic pump is higher than the load pressure by a predetermined value. a discharge amount control means for controlling the discharge amount of a pump, the first actuator being an actuator for driving an inertial load, and the second actuator being an actuator for driving a normal load. In the apparatus, a second hydraulic pump capable of supplying pressure oil to the second actuator separately from the first hydraulic pump, and communication between discharge pipes of the first and second hydraulic pumps is controlled. an opening/closing means; a first detection means for detecting the drive of the first actuator; an instruction means for setting a rate of increase in the flow rate of the pressure oil supplied to the first actuator; and the first detection means. When the drive of the first actuator is detected, a target operating speed of the displacement variable means of the first hydraulic pump is calculated based on the flow rate increase rate set by the instruction means, and the discharge amount is calculated. control means for controlling the displacement variable means so that its operating speed matches the operating speed target value when controlling the discharge amount of the first hydraulic pump by the control means, and switching the opening/closing means to a cutoff position; A hydraulic drive device for construction machinery characterized by: (4) The hydraulic drive device for construction machinery according to claim 3, further comprising second detection means for detecting the drive of the second actuator, and the control means is configured to control the first and second actuators.
When the driving of the first and second actuators are both detected by the detection means, the operation speed of the displacement variable means is controlled and the opening/closing means is switched to a blocking position. Hydraulic drive of the machine.