JP2008256037A - Electric hydraulic work machine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To miniaturize the structure of a system for load sensing. <P>SOLUTION: This electric hydraulic work machine includes: a fixed displacement hydraulic pump 11 driven by an electric motor 10; a plurality of hydraulic actuators 12 driven by pressurized oil from the hydraulic pump 11; an electric lever device 9 outputting an electric manipulate signal as a driving command of each hydraulic actuator 12 by lever operation; control valves 21, 22 controlling pressurized oil flows from the hydraulic pump 11 to the hydraulic actuators 12 according to the manipulate signal; and an electric motor control means 50 controlling the rotation of the electric motor 10. The electric motor control means 50 has: differential pressure detecting means 31, 32 detecting differential pressure between the discharge pressure Pp of the hydraulic pump 11 and the maximum load pressure PLmax of the plurality of hydraulic actuators 12; a frequency calculating means 52 calculating a target frequency fa of the electric motor 10 so that the differential pressure detected by the differential pressure detecting means 31, 32 is made constant; and driving means 53 to 55 each driving the electric motor 10 at a driving frequency f corresponding to the target frequency fa. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to an electric hydraulic working machine that performs various operations by driving a hydraulic pump with an electric motor.

  Conventionally, in a hydraulic excavator that performs operations by combining multiple hydraulic actuators, the displacement of the pump is controlled so that the differential pressure between the pump pressure of the variable displacement hydraulic pump and the maximum load pressure of the hydraulic actuator becomes the set value. A load sensing type hydraulic control device is known (see, for example, Patent Document 1).

JP-A-7-229169

  However, when the regulator is adjusted according to the differential pressure between the pump pressure and the maximum load pressure of the actuator as in the device described in Patent Document 1, and the pump displacement volume is controlled, a large number of hydraulic pipes must be routed. There is a tendency for the apparatus to become larger. Therefore, it is difficult to apply to a mini excavator or the like that requires space saving. Further, in the load sensing type hydraulic control apparatus using hydraulic pressure such as the apparatus described in Patent Document 1, tuning of various hydraulic devices is necessary to make the apparatus function normally, which is troublesome.

An electric hydraulic working machine according to the present invention includes a fixed displacement hydraulic pump driven by an electric motor, at least two first and second hydraulic actuators driven by pressure oil from the hydraulic pump, and a lever operation. An electric lever device that outputs an electrical operation signal that is a drive command for the first and second hydraulic actuators, and controls the flow of pressure oil from the hydraulic pump to the first and second hydraulic actuators according to the operation signal, respectively. First and second control valves, and electric motor control means for controlling the rotation of the electric motor. The electric motor control means includes a difference between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second hydraulic actuators. A differential pressure detecting means for detecting pressure, a frequency calculating means for calculating a target frequency of the electric motor so that the differential pressure detected by the differential pressure detecting means is constant, and a target And having a driving means for driving the motor at a drive frequency corresponding to the wave number.
An upper limit frequency and a lower limit frequency for the motor to output a predetermined torque are set in advance, and when the target frequency is not less than the lower limit frequency and not more than the upper limit frequency, the motor is driven at the target frequency, and when the target frequency is greater than the upper limit frequency, or When smaller than the lower limit frequency, the electric motor can be driven at the upper limit frequency or the lower limit frequency.
Saturation detection means for detecting a saturation state in which the required flow rates of the first and second hydraulic actuators exceed the maximum discharge flow rate of the hydraulic pump, and when the saturation detection state is detected by the saturation detection means, Control valve control means for controlling the first and second control valves so as to limit the supply of pressure oil to the first and second hydraulic actuators at the same rate may be further provided.
In this case, the saturation state can be determined when the target frequency exceeds the upper limit frequency.
A pressure compensation valve that makes the differential pressure across the first and second control valves constant regardless of the load pressure may be further provided.

  According to the present invention, the rotation of the electric motor is controlled so that the differential pressure between the pump discharge pressure and the maximum load pressure of the hydraulic actuator becomes a predetermined value. Applicable.

-First embodiment-
Hereinafter, a first embodiment of an electric hydraulic working machine according to the present invention will be described with reference to FIGS.
FIG. 1 is a side view of a hydraulic excavator which is an example of an electric hydraulic working machine according to the present invention, and particularly shows a mini excavator capable of small turning. The hydraulic excavator includes a traveling body 1, a revolving body 2 that is turnably mounted on the traveling body 1 via a turning device 3, a boom 5, an arm 6, and a bucket 7 that are rotatably attached to the revolving body 2. And a work front 4.

  The boom 5 is rotatably supported by the boom cylinder 5a, the arm 6 is rotatably supported by the arm cylinder 6a, and the bucket 7 is rotatably supported by the bucket cylinder 7a. These cylinders 5 a to 7 a are operated by an operation lever 9 provided in the cab 8. Although the bucket 7 is attached to the tip of the arm, other work attachments can be attached instead of the bucket 7.

  FIG. 2 is a diagram illustrating a system configuration of the excavator according to the first embodiment. 2 includes an electric motor 10, a fixed displacement hydraulic pump 11 driven by the electric motor 10, and a plurality of hydraulic actuators 12 driven by pressure oil from the hydraulic pump 11 (boom cylinder 5a and arm cylinder 6a as an example). And a directional control valve unit 20 that controls the flow of pressure oil from the hydraulic pump 11 to each hydraulic actuator 12, and a control unit 50 that controls the rotation of the electric motor 10.

  The direction control valve unit 20 includes a direction control valve 21 that controls the flow of pressure oil to the boom cylinder 5a, and a direction control valve 22 that controls the flow of pressure oil to the arm cylinder 6a. The direction control valves 21 and 22 are arranged in parallel with each other, and the pressure oil from the hydraulic pump 11 is branched and guided to the direction control valves 21 and 22, respectively. The direction control valves 21 and 22 are switched according to the operation amount of the operation lever 9 (FIG. 1) provided corresponding to each hydraulic actuator 12. The operation lever 9 is an electric lever that outputs an electric operation signal (lever signal) corresponding to the operation amount, and the direction control valves 21 and 22 are electromagnetic proportional valves that are switched by a control signal corresponding to the lever signal. .

  Pressure compensation valves 23 and 24 are provided in the pressure oil supply paths for leading the pump discharge oil to the direction control valves 21 and 22, respectively. The set pressures of the springs 23a and 24a of the pressure compensation valves 23 and 24 are set to a predetermined value Ps, respectively, and the pressure compensation valves 23 and 24 are set so that the differential pressure across the directional control valves 21 and 22 becomes the predetermined value Ps. Operate. As a result, the differential pressure across the direction control valves 21 and 22 is kept constant regardless of changes in the boom cylinder pressure PL1 and the arm cylinder pressure PL2, and the pump flow rate can be divided according to the opening area ratio of the direction control valves 21 and 22. .

  The discharge pressure Pp of the hydraulic pump 11 is detected by the pressure sensor 31. Among the load pressures acting on the actuator 12, the maximum load pressure PLmax is selected by the shuttle valves 25 and 26 and detected by the pressure sensor 32. The pressure signals VP and VL obtained by the pressure sensors 31 and 32 are input to the control unit 50. Reference numeral 27 denotes an unload valve for discharging the surplus flow rate from the hydraulic pump 10 to the hydraulic oil tank when the actuator 12 is not operating and no load, and 28 is a main relief valve for circuit protection. .

  The control unit 50 calculates the pump discharge pressure Pp and the maximum load pressure PLmax from the input unit 51 for inputting the pressure signals VP and VL, and the pressure signals VP and VL, and the difference between the pump discharge pressure Pp and the maximum load pressure PLmax. A calculation unit 52 that calculates a target rotational speed (target frequency fa) of the electric motor 10 for setting the pressure to the target differential pressure Ps, and an output unit 53 that outputs a command signal V0 so as to drive the electric motor 10 at the target rotational speed. The inverter 54 converts the electric power from the external power supply 55 into a driving frequency f corresponding to the command signal V0 and drives the electric motor 10.

  In the present embodiment, the electric motor 10 is controlled so that the differential pressure (Pp−PLmax) between the pump discharge pressure Pp and the maximum load pressure PLmax becomes the target differential pressure Ps. That is, load sensing is performed by a signal from the control unit 50. Here, a deviation between (Pp−PLmax) and the target differential pressure Ps is defined as a correction pressure ΔP (= (Pp−PLmax) −Ps). FIG. 3 is a diagram showing the relationship between the correction pressure ΔP and the target incremental frequency Δf, and this relationship is stored in advance in a memory in the control unit. According to FIG. 3, Δf = 0 when ΔP = 0, and Δf becomes smaller than 0 as ΔP becomes larger than 0, and Δf becomes larger than 0 as ΔP becomes smaller than 0. In the present embodiment, the drive frequency f of the electric motor 10 is corrected as will be described later using the relationship of FIG.

  FIG. 4 is a block diagram showing a detailed configuration of the control unit 50 according to the first embodiment. The pressure signals VP and VL from the pressure sensors 31 and 32 are input to the pressure converters 35 and 36, and converted into the pump discharge pressure Pp and the maximum load pressure PLmax by the pressure converters 35 and 36, respectively. The subtractor 37 subtracts the maximum load pressure PLmax from the pump discharge pressure Pp to obtain a differential pressure (Pp−PLmax). The subtracter 38 subtracts the target differential pressure Ps from the differential pressure (Pp−PLmax) to calculate a correction pressure ΔP.

  The Δf calculator 39 calculates the target incremental frequency Δf from the relationship of FIG. 3 using the correction pressure ΔP as a control parameter. The adder 40 adds the target increment frequency Δf to the current drive frequency f and calculates the target frequency fa. The output unit 53 calculates the drive frequency f from the target frequency fa based on the predetermined characteristics shown in FIG. 5 and outputs a command signal V0 corresponding to the drive frequency f to the inverter 54.

  FIG. 5 is a diagram illustrating the relationship between the target frequency fa and the drive frequency f. In a range where the target frequency fa is not less than the lower limit frequency fmin and not more than the upper limit frequency fmax (fmin ≦ fa ≦ fmax), the drive frequency f is equal to the target frequency fa. When the target frequency fa is smaller than the lower limit frequency fmin, the drive frequency f becomes the lower limit frequency fmin, and when the target frequency fa is larger than the upper limit frequency fmax, the drive frequency f becomes the upper limit frequency fmax. That is, the upper limit value and the lower limit value of the drive frequency f are limited to the upper limit frequency fmax and the lower limit frequency fmin, respectively. The upper limit frequency fmax is set to the base frequency f0 of the electric motor 10, and the lower limit frequency fmin is set to a frequency corresponding to the minimum rotational speed for ensuring the standby flow rate when the unload valve 29 is operated.

  The inverter 54 converts the electric power from the external power supply 55 into a drive frequency f corresponding to the command signal V0, and drives the electric motor 10 at the drive frequency f. FIG. 6 is a diagram showing the relationship between the command signal V0 and the drive frequency f. In FIG. 6, the drive frequency f increases as the command signal V0 increases, and the drive frequency f is controlled according to the command signal V0 based on the characteristics of FIG. Thereby, the electric motor 10 is controlled to the target frequency.

  FIG. 7 is a diagram showing the relationship between the drive frequency f and the output torque T of the electric motor 10. Here, the upper limit frequency fmax is the base frequency f0, and the output torque T at that time is the rated torque Ta. The electric motor 10 is controlled to output a constant rated torque Ta from the lower limit frequency fmin to the upper limit frequency fmax by vector control. The rated torque Ta in this case is set larger than the maximum pump input torque Tmax determined by the displacement volume q of the hydraulic pump 11 and the relief pressure Pr of the main relief valve 28. As a result, the output torque T of the electric motor 10 becomes larger than the maximum pump input torque Tmax in the entire range of the driving frequency f (fmin ≦ f ≦ fmax), and stable speed control of the electric motor 10 is possible.

  The hydraulic excavator according to the present embodiment is used at a work site where exhaust gas and noise are particularly problematic, for example, an indoor site. Power is supplied to the control unit 50 from the external power supply 55 via a cable. The cable is suspended from, for example, an indoor ceiling so that the length does not become insufficient due to a turning operation or the like, and is attached to the control unit 50 on or around the turning center.

  When a drive command for the hydraulic actuator 12 (for example, the boom cylinder 5a and the arm cylinder 6a) is input by operating the operation lever 9, the direction control valves 21 and 22 are driven according to the lever operation amount, and the hydraulic pump 11 supplies each cylinder. The flow of pressure oil to 5a, 6a is controlled. Here, since the directional control valve unit 20 is provided with the pressure compensation valves 23 and 24, even if the load pressures PL1 and PL2 fluctuate, the cylinders 5a and 6a correspond to the operation amount of the operation lever 9. The pump discharge amount can be distributed, and the cylinders 5a and 6a can be appropriately combined.

  The control unit 50 calculates a drive frequency f for setting the differential pressure between the pump discharge pressure Pp and the maximum load pressure PLmax as the target differential pressure Ps, and controls the external power source 55 with the inverter 54 to control the electric motor 10 with the drive frequency f. Drive. For example, when the maximum load pressure PLmax increases during work, (Pp−PLmax) decreases, so the correction pressure ΔP in FIG. 3 becomes negative, and a positive target incremental frequency Δf is output from the Δf calculator 39. As a result, the drive frequency f increases, the rotational speed of the electric motor 10 increases, and the pump discharge rate increases. As a result, the pump discharge pressure Pp increases, the differential pressure between the maximum load pressure PLmax and the pump discharge pressure Pp is controlled to the target differential pressure Ps, and load sensing type hydraulic control can be performed by speed control of the electric motor 10.

According to 1st Embodiment, there can exist the following effects.
(1) Since the speed of the electric motor 10 for driving the hydraulic pump 11 is controlled to set the differential pressure between the pump discharge pressure Pp and the maximum load pressure PLmax to the target differential pressure Ps, the hydraulic pump 11 needs to be configured as a variable displacement pump. This eliminates the need for hydraulic piping for changing the displacement of the pump. As a result, the entire load sensing system can be reduced in size and can be easily applied to a mini excavator. Further, since load sensing is performed by controlling the speed of the electric motor 10, the trouble of tuning the hydraulic equipment can be saved. Since a variable displacement pump is not required, the entire system can be constructed at low cost.
(2) Since load sensing is performed by speed control of the electric motor 10, the number of rotations of the electric motor can be minimized. As a result, for example, energy loss and heat generation can be suppressed as compared to a case where load sensing is performed by changing the pump capacity while rotating the motor 10 at a constant base frequency f0.
(3) An upper limit frequency fmax and a lower limit frequency fmin are set for the drive frequency f of the motor 10, and the motor 10 is driven at a drive frequency f between the upper limit frequency fmax and the lower limit frequency fmin so that the motor 10 always has the rated torque Ta. Since output is performed, stable speed control of the electric motor 10 is possible.
(4) Since the pressure compensation valves 23 and 24 are provided in the directional control valve unit 20 so that the differential pressure across the directional control valves 21 and 22 is constant regardless of the load, the hydraulic actuator 12 is performing load sensing. Can be optimally performed.

-Second Embodiment-
A second embodiment of the electric hydraulic working machine according to the present invention will be described with reference to FIGS.
In the second embodiment, a saturation state in which the pump discharge flow rate is smaller than the required flow rate of the actuator 12 is detected, and a countermeasure against saturation is taken by controlling the direction control valve. In the following description, differences from the first embodiment will be mainly described.

  FIG. 8 is a diagram illustrating a system configuration of a hydraulic excavator according to the second embodiment. In addition, the same code | symbol is attached | subjected to the location same as FIG. The control unit 50 of the second embodiment includes an electric motor control unit 50a for controlling the electric motor 10, a directional control valve 21 for driving the hydraulic actuator 12 (here, the boom cylinder 5a and the arm cylinder 6a), And a control valve control unit 50 b for controlling the control unit 22. The motor control unit 50a has the same configuration as the control unit 50 of FIG.

  The operation lever 9 can be operated from the neutral state in the A direction and the B direction, and the operation amount of the operation lever 9 is detected by the potentiometer 9a. The control valve control unit 50b includes an input unit 61 that inputs a lever signal Lv, a calculation unit 62 that calculates a control signal (drive current) I output to the direction control valves 21 and 22 (electromagnetic proportional valves), and a direction control valve. 21 and 22 have an output unit 63 for outputting a control signal I.

  FIG. 9 is a block diagram showing a detailed configuration of the control unit 50 according to the second embodiment. In addition, the same code | symbol is attached | subjected to the location same as FIG. The operation levers 9A and 9B are operation levers 9 for inputting operation commands for the direction control valves 21 and 22, respectively. Lever signals Lv corresponding to the operation amounts of the operation levers 9A and 9B are input to the input units 61A and 61B of the control valve control unit 50b, respectively. In the signal calculators 64A and 64B, the control signal I corresponding to the input lever signal Lv is calculated based on the predetermined characteristics shown in FIG.

  FIG. 11 is a diagram illustrating the relationship between the lever signal Lv and the control signal I. As shown in FIG. 11, when the control lever 9 is neutral, the lever signal is Lv0. A dead zone is provided near the lever neutral, and the control signal I is zero. When the operating lever 9 is operated from the neutral position in the direction A in FIG. 8, the lever signal Lv increases, and when the operating lever 9 is operated in the direction B, the lever signal Lv decreases. When the operation lever 9 is operated by a predetermined amount or more from the neutral position in the A direction and the B direction, the control signal I increases as the operation amount increases, and when the lever operation amount in the A direction and the B direction is maximum, the control signal I is the maximum Imax.

  The target frequency fa calculated by the motor control unit 50a is input to the gain calculator 65. The gain calculator 65 calculates a correction gain K based on the predetermined characteristics shown in FIG. FIG. 10 is a diagram showing the relationship between the target frequency fa and the correction gain K. As shown in FIG. As shown in FIG. 10, when the target frequency fa is equal to or lower than the upper limit frequency fmax, the correction gain K is 1, and when the target frequency fa exceeds the upper limit frequency fmax, the correction gain K is determined according to a predetermined arithmetic expression (for example, fmax / fa). Decrease gradually. Here, since the drive frequency f of the electric motor 10 is limited so as not to exceed the upper limit frequency fmax (FIG. 5), the saturation state is entered when the target frequency fa exceeds the upper limit frequency fmax, and correction is performed when this saturation state is detected. The gain K is made smaller than 1.

  Multipliers 66A and 66B respectively correct the control signal I by multiplying the control signal I calculated by the signal calculators 64A and 64B by the correction gain K. The output parts 63A and 63B output the corrected control signal I (= K · I) to the solenoids 21a, 21b, 22a and 22b of the directional control valves 21 and 22, respectively. In this case, when the operation levers 9A and 9B are operated in the A direction (FIG. 8), the control signal I is output to the solenoids 21a and 22a, and when the operation levers 9A and 9B are operated in the B direction, the control signal I is output to the solenoids 21b and 22b. . As a result, the directional control valves 21 and 22 are switched to the position A side or the position B side.

  In the above configuration, when the target frequency fa exceeds the upper limit frequency fmax, the saturation state is determined, and the correction gain K is made smaller than 1. As a result, the control signal I output to the directional control valves 21 and 22 is uniformly reduced, and the amount of pressure oil supplied to the hydraulic actuators 5a and 6a is reduced at an equal rate. As a result, the work can be performed in a state where the flow rate ratio at the time of the composite operation is kept constant, and a satisfactory countermeasure against saturation is possible. In this case, the overall work speed is slow, but the speed ratio of the hydraulic actuators 5a and 6a does not change, so the composite operability is good. Since the saturation state is determined when the target frequency fa exceeds the upper limit frequency fmax that is the base frequency f0, the saturation state can be easily detected.

  In the above embodiment, the control signal I is multiplied by the correction gain K as a countermeasure against saturation, but the lever signal Lv may be corrected by multiplying the lever signal Lv by the correction gain K. In this case, for example, the neutral lever signal Lv0 may be set to 0, and the lever signal Lv may be positive when the operation lever 9 is operated in the A direction and negative when the operation lever 9 is operated in the B direction. As a result, when the correction gain K is smaller than 1, the control signal I output to the directional control valves 21 and 22 is also reduced, and the amount of pressure oil supplied to the hydraulic actuators 5a and 6a is suppressed as described above.

  In the above-described embodiment, the case where the boom cylinder 5a and the arm cylinder 6a are combined and operated by the pressure oil from the hydraulic pump 11 has been described. However, the types of the first and second hydraulic actuators are not limited thereto. Alternatively, three or more hydraulic actuators may be combined and operated simultaneously. The configuration of the direction control valves 21 and 22 as the first and second control valves is not limited to that described above. Although the lever signal Lv, which is a drive command for the hydraulic actuator 12, is output by the operation lever 9, any configuration of the electric lever device may be used. For example, if the operation lever 9 can be operated in the cross direction, the drive commands for the two hydraulic actuators 12 can be output by one operation lever 9.

  The rotation of the electric motor 10 is controlled by a signal from the control unit 50 (FIG. 2), but the electric motor is set so that the differential pressure between the pump discharge pressure Pp and the maximum load pressure PLmax of the hydraulic actuator 12 becomes the set differential pressure Ps. As long as the target frequency fa of 10 is calculated and the motor 10 is driven at the drive frequency f corresponding to the target frequency fa, the configuration of the motor control means is not limited to that described above. Although the differential pressure between the pump discharge pressure Pp and the maximum load pressure PLmax is detected based on the detection values of the pressure sensors 31, 32, the configuration of the differential pressure detection means is not limited to this. Although the target frequency fa is calculated using the relationship between the correction pressure ΔP and the target incremental frequency Δf (FIG. 3), the configuration of the frequency calculation means is not limited to this. For example, the relationship between the correction pressure ΔP and the target incremental frequency Δf may be approximated by a straight line as shown in FIG. 12, thereby facilitating the calculation of Δf.

  Although the driving frequency f corresponding to the target frequency fa is calculated by the output unit 53 and the electric motor 10 is driven at the driving frequency f by the signal to the inverter 54, the driving means is not limited to this. Instead of converting the target frequency fa to the drive frequency f according to the characteristics of FIG. 5, the target frequency fa may be output as it is as the drive frequency f. Although the saturation state is detected when the target frequency fa exceeds the upper limit frequency fmax, the saturation detection means is not limited to this. In the above embodiment, the control signal I is multiplied by the correction gain K. However, the control valves 21 and 22 are controlled so as to limit the supply of pressure oil to the hydraulic actuators 5a and 6a at the same rate. If there is, the configuration of the control valve control means is not limited to this.

  Although the said embodiment was applied to the hydraulic shovel (FIG. 1), it can apply similarly to the other electric hydraulic working machine which drives a hydraulic pump with an electric motor and works. That is, as long as the features and functions of the present invention can be realized, the present invention is not limited to the electric hydraulic working machine according to the embodiment.

1 is an external side view of a hydraulic excavator that is an example of an electric hydraulic working machine according to the present invention. The figure which shows the system configuration | structure of the hydraulic excavator which concerns on 1st Embodiment. The figure which shows the relationship between correction | amendment pressure and a target incremental frequency. The block diagram which shows the detailed structure of the control unit which concerns on 1st Embodiment. The figure which shows the relationship between a target frequency and a drive frequency. The figure which shows the relationship between a command signal and a drive frequency. The figure which shows the torque characteristic of an electric motor. The figure which shows the system configuration | structure of the hydraulic excavator which concerns on 2nd Embodiment. The block diagram which shows the detailed structure of the control unit which concerns on 2nd Embodiment. The figure which shows the relationship between a target frequency and correction gain. The figure which shows the relationship between a lever signal and a control signal. The figure which shows the modification of FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 9 Operation lever 10 Electric motor 11 Hydraulic pump 12 Hydraulic actuator 21, 22 Directional control valve 31, 32 Pressure sensor 50 Control unit 50a Electric motor control part 50b Control valve control part 52, 62 Calculation part 53, 63 Output part 54 Inverter 55 External power supply

Claims (5)

A fixed displacement hydraulic pump driven by an electric motor;
At least two first and second hydraulic actuators driven by pressure oil from the hydraulic pump;
An electric lever device that outputs an electric operation signal that is a drive command of the first and second hydraulic actuators by lever operation;
First and second control valves that respectively control the flow of pressure oil from the hydraulic pump to the first and second hydraulic actuators in response to the operation signal;
Electric motor control means for controlling the rotation of the electric motor,
The motor control means includes
Differential pressure detection means for detecting a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second hydraulic actuators;
Frequency calculating means for calculating a target frequency of the electric motor so that the differential pressure detected by the differential pressure detecting means is constant;
An electric hydraulic working machine comprising driving means for driving the electric motor at a driving frequency corresponding to the target frequency. The electric hydraulic working machine according to claim 1,
In the motor control means, an upper limit frequency and a lower limit frequency for the motor to output a predetermined torque are preset,
The driving means drives the motor at the target frequency when the target frequency is not less than the lower limit frequency and not more than the upper limit frequency, and when the target frequency is greater than the upper limit frequency or less than the lower limit frequency, An electric hydraulic working machine that drives an electric motor at the upper limit frequency or the lower limit frequency. The electric hydraulic working machine according to claim 2,
Saturation detection means for detecting a saturation state in which required flow rates of the first and second hydraulic actuators exceed a maximum discharge flow rate of the hydraulic pump;
When the saturation state is detected by the saturation detection means, the first and second control valves are controlled so as to limit the supply of pressure oil to the first and second hydraulic actuators by lever operation at the same rate. An electric hydraulic working machine further comprising control valve control means for controlling. In the electric hydraulic working machine according to claim 3,
The saturation detection means determines that the saturation state is reached when the target frequency exceeds the upper limit frequency. In the electric hydraulic working machine according to any one of claims 1 to 4,
An electric hydraulic working machine, further comprising a pressure compensation valve that makes the differential pressure across the first and second control valves constant regardless of load pressure.

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