JPH0389003A - Inertial hydraulic drive - Google Patents

Abstract

PURPOSE:To prevent excessive power distribution to an actuator and prevent the deterioration of life by controlling a differential pressure change control means such that the differential pressure in front and behind a throttle of a changeover valve is changed with predetermined changing circumstances. CONSTITUTION:A rotating motor 10 and a boom cylinder 15 are controlled by means of changeover valve mechanisms 11 and 16 composed of overload releaf valves 13a, 13b, 18a, 18b, flow control valves 12, 17 and adjusting valves 14, 19. Signals from operation command signal generators 10L, 15L, are input into a control device according to operation rate of an operation lever. An operation signal is output to a solenoid proportional control valve 23 such that differential pressure in front end behind the flow control vale 2 is changed based on a predetermined characteristic, in order to control the adjusting valve 14. It is therefore possible to prevent excessive power distribution to the rotating motor 10, and thereby prevent the deterioration of life of the mechanism.

Description

DETAILED DESCRIPTION OF THE INVENTION [Field of Industrial Application] The present invention relates to a hydraulic drive device for an inertial body that uses pressure oil to drive an actuator that drives the inertial body.

[Prior Art] Machines that use pressure oil to drive a member to perform a desired operation are used in many fields.

One example of such a machine is a hydraulic excavator. A hydraulic excavator is equipped with one or more hydraulic pumps, and the pressure oil discharged from the hydraulic pumps is used to control the boom cylinder, arm cylinder,
Actuators such as the packet cylinder, swing motor, and left and right travel motors are driven. A flow control valve is provided between each of these seven actuators and the hydraulic pump, and the flow control valve can be controlled by the operator arbitrarily operating the control lever of each actuator provided in the driver's seat of the hydraulic excavator. is operated accordingly, controlling the flow direction and flow rate of pressure oil from the hydraulic pump to the actuator, which in turn controls the operation of the actuator, thereby remotely performing the intended work of the hydraulic excavator.

Various means have been proposed for drive control of such actuators, and one of them is, for example, Japanese Patent Laid-Open No. 60
There is a load sensing system disclosed in Japanese Patent No.-11706. In this load sensing system, during drive control of the actuator, the differential pressure between the hydraulic pressure on the hydraulic pump side of the flow control valve and the hydraulic pressure (load pressure) on the actuator side (differential pressure generated by the metering orifice of the flow control valve) is constantly maintained. , a means for controlling a hydraulic pump (variable displacement hydraulic pump) to a certain predetermined value. More specifically, in a state where the variable displacement hydraulic pump discharges a constant flow rate determined by the tilt angle of its swash plate, etc., and the actuator is being driven by this, the operating lever is pressed to increase the speed of the actuator. When is operated, the opening area of the flow control valve increases accordingly. Therefore, the differential pressure inevitably decreases and a deviation from the predetermined value occurs. In accordance with this deviation, a command is output to the variable displacement hydraulic pump to increase its tilt angle to increase the discharge flow rate of pressure oil.

As a result, the amount of oil passing through the flow control valve increases and the speed of the actuator is increased, and at the same time, the differential pressure also increases and returns to the predetermined value. In this way, by controlling the variable displacement hydraulic pump so as to always maintain the differential pressure at a predetermined value, the actuator can be driven and controlled in accordance with the operation of the operating lever.

[Problems to be Solved by the Invention] In the load sensing system, the flow rate to the actuator is determined only by the opening degree of the flow control valve within the torque range limited by the input torque control means of the variable displacement hydraulic pump. This is an extremely excellent means for controlling the speed of an actuator. However, such a load sensing system has disadvantages when driving an inertial human load (hereinafter simply referred to as an inertial body).

This will be explained below using an example of a swing motor of a hydraulic excavator. A swing motor of a hydraulic excavator is an actuator that rotates an upper revolving body to which a front mechanism is attached and a prime mover, a hydraulic motor, a driver's cab, and other devices are installed, that is, an actuator that drives an inertial body. Therefore, a large pressure is required to start the swing motor, and for this reason, the operator of the hydraulic excavator usually maximizes the amount of operation of the operating lever when starting the swing motor. By the way, in the load sensing system, when the operating amount of the operating lever is maximized, the opening degree of the flow control valve also becomes the maximum opening degree, and the variable displacement hydraulic pump tries to supply the flow rate according to this maximum opening degree. Therefore, when the swing motor is accelerated, a power determined by the flow rate corresponding to the maximum opening and the relief valve setting pressure is applied to the swing motor, and the swing motor is accelerated by this power.

For this reason, excessive power is applied to a power transmission mechanism such as a gear interposed between the upper revolving structure and the rotation motor, which has an adverse effect on the power transmission mechanism, such as shortening its lifespan.

Furthermore, surplus flow other than the flow actually flowing into the swing motor flows out into the tank via the relief valve, causing power loss.

To prevent such inconvenience in the load sensing system, it is possible to limit the input torque of the variable displacement hydraulic pump. If this is done, the power distributed to each actuator will be reduced and the working capacity will be reduced, so it cannot be adopted at all.

The above-mentioned inconvenience is a problem that always occurs not only in the swing motor of a hydraulic excavator but also in hydraulic actuators that drive inertia bodies in various hydraulic machines.

The purpose of the present invention is to solve the problems in the above-mentioned prior art,
In hydraulic equipment that uses a load sensing system, the hydraulic pressure of the inertial body can prevent excessive power from being distributed to the actuator that drives the inertial body, and can also prevent shortening of the life of the mechanism. To provide a thigh motion device.

[Means for Solving the Problems] In order to achieve the above object, the present invention provides an inertial body, a hydraulic actuator that drives the inertial body, a hydraulic pump that supplies pressure oil to the hydraulic actuator, and the hydraulic actuator. In a hydraulic drive device comprising a switching valve that controls the drive of the switching valve, and differential pressure control means that controls the differential pressure across the flow rate control throttle of the switching valve, when a switching command to the switching valve is output, the difference A differential pressure change control means is provided for driving and controlling the pressure control means so that the differential pressure across the throttle of the switching valve changes in a predetermined manner, and the power applied to the hydraulic actuator is controlled to a predetermined value or less. It is characterized by being limited to.

[Operation] In a hydraulic actuator that drives an inertial body, when a switching command is output to a switching valve that controls its drive, the differential pressure across the throttle of the switching valve is changed by the differential pressure change control means. By driving, it changes in a predetermined manner, and finally reaches a predetermined set value.

[Example] Hereinafter, the present invention will be explained based on illustrated embodiments.

FIG. 1 is a hydraulic circuit diagram of a hydraulic movement device for a hydraulic excavator according to an embodiment of the present invention. In FIG. 9, reference numeral 1 indicates a pressure oil supply mechanism. This pressure oil supply mechanism 1 includes a variable displacement hydraulic pump 2,
3. Variable displacement mechanisms (represented by swash plates) 2a, 3a, hydraulic cylinders 4 that drive each swash plate 2a, 3a, control valve 5 that controls input torque, and control valve that controls the load sensing system. It consists of 6.

7 is a pilot pump, and 8 is a relief valve that defines the maximum pressure in the circuit of the pilot pump 7.

10 is a swing motor of the hydraulic excavator, IOL is an operation command signal generator that outputs a signal according to the amount of operation of a control lever (not shown) of the swing motor 10, and 11 is a switching valve mechanism of the swing motor 10. The switching valve mechanism 11 includes overload relief valves 13a, 13b, a flow control valve 12, and an adjustment valve 14. Reference numeral 15 indicates a boom cylinder of the hydraulic excavator, ↓5L indicates an operation command signal generator that outputs a signal according to the operation amount of an operation lever (not shown) of the boom cylinder 15, and 16 indicates a switching valve mechanism of the boom cylinder 15. The switching valve mechanism 16, like the switching valve mechanism 11, includes a flow control valve 17, overflow relief valves 18a, 18b, and a regulating valve 19. 20 is the swing motor 1
0 and the load pressure of the boom cylinder 15, and 21 is a relief valve that defines the maximum pressure of the hydraulic circuits of the variable displacement hydraulic pumps 2 and 3. Reference numeral 22 denotes a control device which inputs signals from the operation command signal generators 10L and 15L and controls the opening degrees of the flow rate control valves 12 and 17 based on the signals, and is composed of a microcomputer. Reference numerals 23 and 24 are electromagnetic proportional pressure reducing valves, and the control pressures outputted from these are guided to regulating valves 14 and 19, respectively.

Next, a flowchart showing the operation of this embodiment in FIG.
This will be explained with reference to the differential pressure change target value characteristic diagram shown in FIG. 3 and the drive pressure characteristic diagram shown in FIG. 4. Swing motor 1
When the operation lever 0 is operated, a signal is output from the operation command signal generator 10L according to the amount of operation, and based on this signal, the control device 22 drives the flow rate control valve 12 to a predetermined opening degree. The operation of this control device 22 will be described later.

When the flow control valve 2 is opened, pressure oil from the variable displacement hydraulic pumps 2 and 3 is supplied to the swing motor 10, and the swing motor 10 is operated. The boom cylinder 15 is also driven by a similar operation.

In driving such a hydraulic actuator, the control valve 5 controls the input torque by introducing the discharge pressure of the variable displacement hydraulic pumps 2 and 3 to limit the input torque from exceeding a certain set value. 6 is a variable displacement hydraulic pump 2, 3
The discharge pressure and the maximum load pressure from the shuttle valve 20 are derived to perform load sensing control. Furthermore, the regulating valve 14 has the function of adjusting the bias in the flow rate of pressure oil due to the compensation of load pressure when a complex operation is performed, and also reduces the differential pressure before and after the flow control throttle of the upper flow control valve 2 by electromagnetic proportional reduction. It also has a function of controlling according to the control pressure guided from the valve 23.

Here, the operation of the control device 22 when command signals are output from the operation command signal generators 10L and L5L will be described. The control device 22 first reads these command signals (
Step S□) in the flowchart shown in FIG. Then, an operation command signal generation 115 is generated for the flow rate control valve 17.
Outputs an active signal according to the command signal output from L,
The flow rate control valve 17 is set to an opening degree according to the thigh motion signal y. As a result, the boom cylinder 15 is moved at a speed corresponding to its opening degree.

On the other hand, the swing motor 10 is controlled by the @9j device 22 to avoid applying excessive power as described above. This control is based on the following idea. That is, now, the moment of inertia of the inertial body is J, the rotational angular velocity of the upper rotating body of the hydraulic excavator is ω, and the driving pressure is P.
, when the capacity of the swing motor 10 is D and the flow rate to the swing motor 10 is Q, the following equation holds true.

Here, since ω=Q/D, it becomes. Equation (2) shows that the pressure P applied to the swing motor 10 can be controlled by adjusting the temporal change in the flow rate Q (dQ/dt).

By the way, the flow rate Q to the swing motor 10 is such that the differential pressure is Δ
When P and the coefficient are G, it is expressed as Q=G-F (3), so the change in flow ff1Q with respect to time is as shown in the following equation.

As a result, the active pressure P of the swing motor 10 can be controlled by the change in the differential pressure ΔP over time from equations (2) and (4) above.

The control device 22 outputs a predetermined signal that changes with time to the electromagnetic proportional pressure reducing valve 23 to control the regulating valve 14, and the regulating valve 14 controls the flow rate control throttle differential pressure of the flow rate control valve 12. It is something to do. Through such control, the throttle differential pressure is controlled in a time-functional manner, so that the time change in the flow rate supplied to the swing motor 1o is adjusted, thereby controlling the power applied to the swing motor 10. That will happen. Step 8 of the flowchart shown in Figure 2
The process 2 or less is a process for performing such control. Next, this process will be explained.

When the signal from the operation command signal generator 10L is read in step S, control is performed to change the set differential pressure of the regulating valve 14 using this signal as a control start signal. That is, first, the differential pressure change target value is determined from the previous command value of the set differential pressure of the regulating valve 14 (command value obtained in step S3 described later), that is, the increment Δ of the value after a certain period of time, 17〒-) (Procedure S2L FIG. 3 shows the characteristics of the differential pressure change target value with respect to the differential pressure command value. Such characteristics are determined by the control device 22
or stored in a memory device. As is clear from the solid line shown in FIG. 3, the differential pressure change target value is set to be the maximum value when the differential pressure command value is from O to a certain value, and thereafter is set to decrease linearly. This is because this is the first process when starting up the swing motor 10.

The previous differential pressure command value was O.

Next, in order to output the first differential pressure command value to the regulating valve 14, this is calculated as the current differential pressure command value (step S). That is, the current differential pressure command value is obtained by adding (differential pressure change target [XΔT) to the previous differential pressure command value. In this case, the previous differential pressure command value is O1, the differential pressure change target value is the value determined by the process in step S2, and ΔT is a predetermined unit time. The current differential pressure command value obtained in this way is finally compared with a value corresponding to the maximum value of the differential pressure controlled by the regulating valve 14 (maximum differential pressure equivalent value) (step S4). If the current differential pressure command value has not reached the maximum differential pressure equivalent value, the process moves to step S, and the current differential pressure command value is output. Then, the process returns to step S again,
The processing of steps 81 to Ss is repeated.

By repeating the above process, the differential pressure set by the regulating valve 14 gradually increases regardless of the operating amount of the operating lever of the swing motor 10, and the increase is caused by the differential pressure shown in FIG. It will be responsive to change. On the other hand, if it is determined in step S that the current differential pressure command value has reached the maximum differential pressure equivalent value.

Regardless of the current differential pressure command value obtained in step S3, replace it with the maximum differential pressure equivalent value (step S,),
This value is output to the regulating valve 14 (step S).

Through the above control, the active flow rate of the swing motor 10 is controlled by temporal changes in the differential pressure before and after the flow rate control throttle. Here, since the active pressure of the swing motor is proportional to the flow rate to the swing motor 10, the active pressure of the swing motor 10 is ultimately controlled by the differential pressure. FIG. 4 is a characteristic diagram of the driving pressure with respect to the flow rate of the swing motor 10 when the differential pressure command value shown in FIG.
It is clear that the power applied to zero can be controlled so that it does not exceed a predetermined value.

FIG. 5 is a characteristic diagram of temporal changes in the target differential pressure change value Δ(rTj-) of the regulating valve 14, the flow rate Q, and the differential pressure command value 5〒- before and after the flow rate control valve 12 in the above control. A time t is set for . As is clear from the figure, time 1 =
0, the regulating valve 14 is made a large throttle by the output of the electromagnetic proportional pressure reducing valve 23, and the value Δ (11 valves) is large. This value Δ(p〒-) remains a constant value as time passes until time t. That is, the period between time t0 and t1 corresponds to the maximum target value region shown in FIG. 3, and therefore the command value β7 of the differential pressure across the flow rate control valve 12 increases linearly. Correspondingly, the flow rate Q is initially very small despite the flow control valve 12 being open, and increases almost linearly as the value decreases.

Between time t1 and t2, the value Δ(men〒-) is the third
The area is reduced corresponding to the shaded area shown in the figure.

Correspondingly, increases in the flow rate Q and the differential pressure command value F are also suppressed. After time t3, the value Δ(p) is almost 0.
, the flow rate Q and the differential pressure command value F are constant values. That is, the swing motor 10 exits the starting state and enters a steady rotation state.

In this way, in this embodiment, the flow rate to the swing motor is increased by changing the differential pressure command value of the regulating valve at a predetermined rate of change, and a load sensing system is adopted. Also, it is possible to prevent excessive power from being applied to the swing motor, which in turn prevents the life of the transmission mechanism from being shortened, and furthermore prevents power loss caused by relief of surplus flow. I can do it. In addition, hydraulic excavators often perform work in which the boom is raised and the upper rotating structure is rotated at the same time, but in this work, the rotating speed of the upper rotating structure is generally lower than the boom raising speed. I don't like being too early. However, by employing the control means of this embodiment and arbitrarily setting the characteristics of the flow rate change rate, it is possible to appropriately distribute the power, and as a result, both can be activated in a timely manner.

In addition, in the explanation of the above embodiment, the swing motor of a hydraulic excavator was used as an example of the hydraulic actuator, but the invention is not limited to this, and any hydraulic actuator that actively operates an inertial body can be used. Applicable. Furthermore, if the characteristic of the differential pressure change target value shown in FIG. 3 is made hyperbolic as shown by the broken line, better equihorsepower approximation becomes possible. Furthermore, even better characteristics can be obtained by correcting the above characteristics using the moment of inertia and the pump rotation speed.

[Effects of the Invention] As described above, in the present invention, the differential pressure change control means changes the set differential pressure of the differential pressure control means based on predetermined characteristics. It is possible to prevent excessive power from being distributed to the hydraulic actuator that activates the actuator, and in turn, it is possible to prevent a reduction in the life of the mechanism. Further, when a hydraulic actuator other than the above hydraulic actuator is provided and a combined operation with the hydraulic actuator is performed, each hydraulic actuator can be operated appropriately.

[Brief explanation of drawings]

FIG. 1 is a hydraulic circuit diagram of a hydraulic active device of a hydraulic excavator according to an embodiment of the present invention, and FIGS. 2, 3, and 4 are flow charts explaining the operation of the control device shown in FIG. The pressure change target value characteristic diagram and the active pressure characteristic diagram, FIG. 5, are graphs showing changes in values Δ<5〒-), Q, and p. 2, 3... Variable displacement hydraulic pump, 5, 6...
...control valve, ↓O...swivel motor, IOL,
15L...Operation command signal generator, 12.17.
...Flow control valve, 14.19 ...Adjustment valve,
15... Boom cylinder, 20... Shuttle valve, 22... Control device, 23° 24...
...Solenoid proportional pressure reducing valve. Fig. Fig. Differential pressure indicator value (MuT) Fig.

Claims (1)

[Claims] An inertial body, a hydraulic actuator that drives it, a hydraulic pump that supplies pressure oil to the hydraulic actuator, a switching valve that controls the drive of the hydraulic actuator, and a differential that controls the differential pressure across the throttle of the switching valve. In the hydraulic drive device, the differential pressure control means is configured to change the differential pressure across the throttle of the switching valve in a predetermined manner when a switching command to the switching valve is output. 1. A hydraulic drive device for an inertial body, characterized in that a differential pressure change control means is provided for drive control.