JP2860158B2 - Hydraulic drive for civil and construction machinery - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a hydraulic drive device for civil engineering and construction machinery such as a hydraulic shovel, and more particularly, to a pressure compensating valve for controlling a differential pressure between a flow control valve for controlling the drive of an actuator and a flow control valve for controlling the actuator. Hydraulic drive device for civil engineering and construction machinery.

BACKGROUND ART Hydraulic excavators include hydraulic drive devices used in civil engineering and construction machines, such as a hydraulic pump discharge output, that is, a discharge flow rate of a hydraulic pump such that a pump pressure is higher than a load pressure of an actuator by a constant value. That is, there is a system called a load sensing system that controls a pump flow rate and discharges only a flow rate necessary for driving an actuator from a hydraulic pump. This load sensing system
For example, as described in Japanese Patent Application Laid-Open No. 60-11706, a working cylinder that controls the displacement of a hydraulic pump, and operates in response to a differential pressure between a pump pressure and a load pressure to control the driving of the working cylinder. A pump regulator for load sensing control (LS control) having a switching valve is provided.
The switching valve is provided with a spring for biasing the switching valve so as to oppose the differential pressure between the pump pressure and the load pressure. The switching valve is provided by a balance between the differential pressure between the pump pressure and the load pressure and the force of the spring. Is operated, and the pump flow rate is controlled such that the differential pressure is maintained at a constant value corresponding to the force of the spring, that is, the target differential pressure.

In the load sensing system, a pressure compensating valve that controls the differential pressure across the flow control valve is arranged upstream of the flow control valve to secure a flow control function for fluctuations in the differential pressure between the pump pressure and the load pressure. It is common.

The pressure compensating valve is generally slidably disposed within the valve housing and has a valve spool having a flow control providing a variable restriction, and a pressure spool formed within the valve housing and located at opposite ends of the valve spool. First and second control chambers facing each other, an inlet pressure of the flow control valve is guided to the first control chamber in the valve closing direction, and an actuator is provided to the second control chamber in the valve opening direction. Load pressure (flow control valve outlet pressure)
Is led. A spring for biasing the valve spool in the valve opening direction is disposed in the second control chamber, and the spring provides a target value for pressure compensation. The spring of the pump regulator and the spring of the pressure compensating valve are set so that the target differential pressure of the LS control is larger than the target value of the compensation differential pressure. And the first and second
The valve spool is actuated by the differential pressure between the inlet pressure of the flow control valve guided to the control chamber and the load pressure of the actuator, that is, the balance between the differential pressure before and after the flow control valve and the force of its spring. Controls the differential pressure between front and rear.

That is, until the differential pressure between the pump pressure controlled by the pump regulator and the load pressure, that is, the LS differential pressure reaches a constant value corresponding to the spring force, that is, the target value of the compensation differential pressure, the valve spool is It is held fully open by the force of the spring,
When the LS differential pressure becomes larger than the target value of the compensation differential pressure, and the differential pressure across the flow control valve attempts to exceed the target value, the valve spool moves in the valve closing direction against the force of the spring, and the flow rate is reduced. The controller is throttled to maintain the differential pressure across the flow control valve at its target value. As a result, the flow rate flowing through the flow control valve, that is, the flow rate supplied to the actuator becomes a flow rate proportional to the opening area of the flow control valve, and stable control of the actuator becomes possible.

Incidentally, this type of pressure compensating valve is, for example, U.S. Pat.
No. 600.

However, the hydraulic drive device employing the conventional pressure compensating valve has the following problems.

In the above-mentioned conventional pressure compensating valve, when the pressure oil flows through the flow control unit of the valve spool, a force acting on the valve spool in the valve closing direction, that is, a flow force is generated by the flow. This flow force is a function of the flow rate and the flow rate flowing through the flow rate control unit. When the flow rate is substantially constant, the flow rate is a function of the differential pressure across the flow rate control unit, and as these increase, the flow force increases. Therefore, when the LS differential pressure increases, the flow force increases, and the set value of the spring, that is, the target value of the compensation differential pressure, decreases due to the increase of the flow force. As a result, the characteristic of the differential pressure across the flow control valve with respect to the LS differential pressure has a negative gradient, and the characteristic of the flow rate passing through the flow control valve with respect to the LS differential pressure also has a negative gradient. Becomes That is, when the flow rate through the flow control valve decreases, the LS differential pressure increases, and when the flow rate increases, the LS differential pressure increases.
The characteristic is such that the differential pressure decreases.

Therefore, in the conventional pressure compensating valve, the LS differential pressure is maintained at the target differential pressure by the control of the pump regulator, and when the valve spool of the pressure compensating valve is in the throttle state, the pump flow rate decreases for some reason, and the flow rate control is performed. If the flow rate through the valve decreases correspondingly, the pressure compensating valve operates to increase the LS differential pressure. On the other hand, when the LS differential pressure increases,
The pump regulator reduces the pump flow rate to maintain the LS differential pressure at the target differential pressure. Therefore, the flow rate through the flow control valve further decreases, and the pressure compensating valve operates to further increase the LS differential pressure. Conversely, the pump flow rate increases,
When the flow rate through the flow control valve increases, the pressure compensating valve operates to decrease the LS differential pressure, and in response to the decrease in the LS differential pressure, the pump regulator holds the LS differential pressure at the target differential pressure. The pump flow rate is increased as much as possible, and as the flow rate through the flow control valve increases, the pressure compensating valve operates to further reduce the LS differential pressure.

As described above, in the conventional hydraulic drive device including the pressure compensating valve, the pump flow rate is controlled to be unilaterally decreased or increased due to the influence of the flow force, and there is a risk that the pump flow rate becomes uncontrollable. There is.

In consideration of such a change in the characteristics of the pressure compensating valve due to the influence of the flow force, a damping means, for example, a throttle is provided in a passage through which the inlet pressure of the pressure compensating valve is led to delay the response to a change in the pump flow rate. However, the provision of such a throttle causes a problem that the response of the pressure compensating valve as the original pressure compensating function is deteriorated.

An object of the present invention is to provide a hydraulic drive device capable of preventing deterioration of control characteristics of a pressure compensating valve due to the influence of a flow force and performing stable pump flow control.

DISCLOSURE OF THE INVENTION In order to achieve the above object, according to the present invention, there is provided a hydraulic pump, an actuator driven by pressure oil discharged from the hydraulic pump, and a flow control disposed between the hydraulic pump and the actuator. A valve, a pressure compensating valve for controlling a differential pressure across the flow control valve, and a pump flow control means for controlling a flow discharged from the hydraulic pump in accordance with a differential pressure between the pump pressure and a load pressure of the actuator. Wherein the pressure compensating valve includes a valve body, and a first control for applying a first control force to the valve body based on a pressure difference between the front and rear of the flow control valve and biasing the valve body in a valve closing direction. Means, and a second control means for applying a predetermined second control force to the valve body and urging the valve body in the valve opening direction, wherein the pressure compensating valve is , The pump pressure and the A third control force that continuously increases as the differential pressure increases based on the differential pressure with respect to the load pressure of the actuator, and applies a third control force to the valve body to urge the valve body in the valve opening direction;
And a hydraulic drive device for a civil engineering / construction machine.

Based on the pressure difference between the pump pressure and the load pressure of the actuator, that is, based on the LS pressure difference, a third control force that continuously increases as the pressure difference increases is applied to the valve body of the pressure compensating valve. By energizing in the valve opening direction, the characteristic of the differential pressure across the flow control valve with respect to the LS differential pressure has a positive slope with the latter increasing as the LS differential pressure increases. Then, the characteristic of the flow rate of the flow control valve with respect to the LS differential pressure also has a positive slope in which the flow rate increases as the LS differential pressure increases. For this reason, even if the pump flow rate fluctuates, the LS
The differential pressure is controlled to return to the target differential pressure, thereby enabling stable control of the pump regulator.

Preferably, the third control means includes a first pressure receiving means to which the pump pressure is guided and urges the valve body in a valve opening direction. This first pressure receiving means may be formed inside the valve body, or may be formed outside the valve body.

Preferably, the first control means includes: a first control chamber into which an inlet pressure of the flow control valve is led;
A first pressure receiving portion that is disposed in the control chamber of the first pressure chamber and biases the valve body in the valve closing direction; a second control chamber into which the load pressure is guided; and a second pressure chamber that is disposed in the second control chamber and includes the valve A second pressure receiving portion that urges the valve body in the valve opening direction; and a third pressure receiving portion that urges the valve member in the valve opening direction when the pump pressure acts on the second pressure receiving portion. The sum of the pressure receiving area of the second pressure receiving section and the pressure receiving area of the third pressure receiving section is substantially equal to the pressure receiving area of the first pressure receiving section.

Preferably, the third control means is formed in the valve body in the axial direction, one end is closed, and the other end is the second control means.
A piston rod that is slidably inserted into the cylinder chamber and protrudes out of the valve body from the other end; and a pump that is formed in the valve body and is provided in the cylinder chamber. A third pressure receiving portion formed at a closed end of the cylinder chamber. The valve body includes a large-diameter portion at one end thereof, and the first pressure-receiving portion is formed on an end face of the large-diameter portion, and the third control means controls the first pressure-receiving portion of the large-diameter portion. And the third pressure receiving portion may be formed on the annular end surface.

The second control means may be a spring or a hydraulic means for generating the second control force by hydraulic pressure. In the latter case, preferably, the hydraulic means includes first hydraulic pressure generating means for generating a constant hydraulic pressure, and second pressure receiving means for guiding the constant hydraulic pressure to urge the valve body in a valve opening direction. And second pressure generating means for generating a variable oil pressure, and third pressure receiving means for guiding the variable oil pressure to bias the valve body in a valve closing direction.

Further, according to the present invention, there is provided a pressure compensating valve for controlling a differential pressure across a flow control valve disposed between a hydraulic pump and an actuator, the spool compensating valve having a spool bore, an inlet recess connected to the hydraulic pump, and A valve housing having an outlet recess connected to the flow control valve; a valve spool slidably disposed in the spool bore for controlling communication between the inlet recess and the outlet recess; Formed in
A first control chamber into which an inlet pressure of the flow control valve is led;
A first pressure receiving portion that is disposed in the first control chamber and urges the valve spool in a valve closing direction; and a second control chamber that is formed in the valve spool and guides a load pressure of the actuator. A second pressure receiving portion disposed in the second control chamber and for urging the valve spool in the valve opening direction; and a second pressure receiving portion for urging the valve spool in the valve opening direction with a predetermined control force to reduce the compensation differential pressure. A pressure compensating valve having means for setting a target value, further comprising: a control means including a third pressure receiving portion that guides the pressure of the inlet recess and urges the valve spool in the valve opening direction. A pressure compensating valve is provided, wherein the sum of the pressure receiving area of the second pressure receiving section and the pressure receiving area of the third pressure receiving section is substantially equal to the pressure receiving area of the first pressure receiving section.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a schematic diagram showing a hydraulic drive device according to a first embodiment of the present invention.

FIG. 2 is a diagram showing characteristics of a spring in the pressure compensating valve.

FIG. 3 is a diagram showing the influence of the flow force on the pressure compensating valve.

FIG. 4 is a diagram showing the effect of the LS differential pressure on the pressure compensating valve.

FIG. 5 is a characteristic diagram showing the relationship between the LS differential pressure obtained in the first embodiment and the differential pressure across the flow control valve.

FIG. 6 is a characteristic diagram showing the relationship between the LS differential pressure obtained in the first embodiment and the flow rate through the flow control valve.

FIG. 7 is a characteristic diagram showing the relationship between the LS differential pressure obtained by the conventional pressure compensating valve and the differential pressure across the flow control valve.

FIG. 8 is a characteristic diagram showing the relationship between the LS differential pressure obtained by the conventional pressure compensating valve and the flow rate through the flow control valve.

FIG. 9 is a schematic view showing a hydraulic drive device according to a second embodiment of the present invention.

FIG. 10 is a schematic diagram showing a hydraulic drive device according to a third embodiment of the present invention.

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, some preferred embodiments of the present invention will be described with reference to the drawings.

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

In FIG. 1, a hydraulic drive device according to the present embodiment includes a variable displacement hydraulic pump 1, an actuator 2 driven by pressurized oil discharged from the hydraulic pump 1, and a hydraulic pump 1 and an actuator 2. Flow control valve 5 which is disposed in the pipelines 3, 4 a and 4 b and controls the driving of the actuator 2.
And the pipeline on the upstream side of the flow control valve 5, that is, the hydraulic pump 1
And a pressure compensating valve 8 arranged in the discharge pipes 6 and 7 for controlling the pressure difference Pz-PLS before and after the flow control valve 5, and the discharge flow rate of the hydraulic pump 1, that is, the pump flow rate There is provided a pump regulator 9 which controls the discharge pressure, that is, a pump pressure Pd and a pressure difference Pd-PS between the load pressure PLS of the actuator 2. Flow control valve 5 and pressure compensating valve 8
A check valve 10 for preventing backflow is disposed in the pipelines 3 and 7 between them, and the inlet pressure Pz of the flow control valve 5 is taken out by a pipeline 11 connected to the pipeline 3, Outlet pressure,
That is, the load pressure PLS of the actuator 2 is controlled by the flow control valve 5.
Is detected by the load line 12 connected to.

The pump regulator 9 is connected to the swash plate 1a of the hydraulic pump 1 and controls the displacement of the hydraulic pump 1 by an actuator 13, and a differential pressure Pd between the pump pressure Pd and the load pressure PLS.
A switching valve 14 that operates in response to the PLS and controls the driving of the actuator 13; The actuator 13 is
A piston 13a having both end faces having different pressure receiving areas, and a double-acting cylinder having a small-diameter cylinder chamber 13b and a large-diameter cylinder chamber 13c located on both end faces of the piston 13a,
The small-diameter cylinder chamber 13b communicates with the discharge pipe 6 of the hydraulic pump 1 via the pipe 15, the large-diameter cylinder chamber 13c communicates with the discharge pipe 6 via the pipe 16, the switching valve 14 and the pipe 17, and The tank 19 is connected to the tank 19 via a pipe 16, a switching valve 14, and a pipe 18.
The switching valve 14 has two opposing driving units 14a and 14b, and one driving unit 14a is loaded with the pump pressure Ps from the pipes 20 and 17 and the other driving unit 14b is loaded with the pump pressure Ps from the load pipe 12. Pressure PL
S is loaded. Further, a spring 14c is disposed on the drive unit 14b side of the switching valve 14.

When the load pressure PLS detected in the load line 12 increases, the switching valve 14 is driven to the left in the drawing to assume the position shown in the drawing, the large-diameter cylinder chamber 13c of the actuator 13 communicates with the discharge line 6, and the piston 13a Piston 1
3a is moved to the left in the figure to increase the amount of tilt of the swash plate 1a, that is, the displacement. As a result, the pump flow increases and the pump pressure Pd increases. When the pump pressure Pd rises, the switching valve 14 is returned to the right in the figure, and the differential pressure Pd-PLS is
When reaching the target value determined by c, the switching valve 14 stops,
The pump flow rate will be constant. Conversely, when the load pressure PLS decreases, the switching valve 14 is driven to the right in the figure, and the large-diameter cylinder chamber 13
c communicates with the tank 19, the piston 13a is moved rightward in the figure, and the tilt amount of the swash plate 1a decreases. As a result, the pump flow rate decreases, and the pump pressure Pd decreases. When the pump pressure Pd decreases, the switching valve 14 is returned to the left in the figure. When the differential pressure Pd-PLS reaches a target value determined by the spring 14c, the switching valve 14c stops, and the pump flow rate becomes constant. Thus, the differential pressure Pd-P
The pump flow rate is controlled so that LS is maintained at the target differential pressure determined by the spring 14c.

The pressure compensating valve 8 has an inlet port 21a and an outlet port 21b, two control ports 21c, 21d, and has a spool housing 22 formed therein, and is slidable in the spool bore 22 in the axial direction. And the valve spool 23 that has been sought. Annular inlet recesses 24 and outlet recesses 25 communicating with the inlet port 21a and the outlet port 21b are formed in the valve housing 21, respectively.
In a, a plurality of notches 26 forming a variable aperture are formed between the inlet recess 24 and the outlet recess 25.

In the valve housing 21, pressure receiving portions 27 and 28 formed by end faces of both ends of the valve spool 23 are located, and two control chambers for urging the valve spool 23 in the valve closing direction and the valve opening direction, respectively.
29, 30 are formed, and these control chambers 29, 30 communicate with two control ports 21c, 21d, respectively. A spring 31 is arranged in the control room 30.

The inlet port 21a is connected to the discharge line 6, and the outlet port 2
1b is connected to the pipeline 7, the control port 21c is connected to the pipeline 11, and the control port 21d is connected to the load pipeline 12.

In the present embodiment, a cylinder chamber 32 having one end closed and the other end opened to the end surface of the pressure receiving portion 28 of the valve spool 23 is formed in the valve spool 23 in the axial direction. A piston rod 33 projecting into 30 is slidably inserted. The valve spool 23 has a cylinder chamber 32.
A passage 34 communicating with the inlet recess 24 is formed in the cylinder chamber.
A pressure receiving portion 35 is formed at the closed end of the portion 32.

The pressure receiving area of the pressure receiving section 27 is Az, and the pressure receiving area of the pressure receiving section 28 is AL.
S, if the pressure receiving area of the pressure receiving unit 35 is Ad, the relationship is Az = ALS + Ad.

In the above configuration, the control chambers 29, 30 and the pressure receiving portions 27, 28 apply a first control force to the valve spool 23 based on the pressure difference Pz-PLS between the flow control valve 5 and the valve spool 23 in the valve closing direction. To provide a first control means for biasing the valve spool 23
And a second control means for applying a second control force based on the spring constant to bias the valve spool 23 in the valve opening direction. The control chamber 29 and the pressure receiving section 27, the cylinder chamber 32 and the pressure receiving section 35 are provided with a differential pressure between the pump pressure and the load pressure controlled by the load sensing control (LS control) by the pump regulator 9, that is, the LS differential pressure Pd-PLS. The third control means that applies a third control force that increases in accordance with the increase of the pressure to the valve spool 23 and urges the valve spool 23 in the valve opening direction is provided.

In the first embodiment configured as described above, when the flow control valve 5 is at the neutral position, the valve spool 23 is moved leftward in the figure by the spring 31, and the pressure compensating valve 8 is in the fully open state. . At this time, the swash plate 1a of the hydraulic pump 1 is held at the minimum tilt position by the pump regulator 9.

In such a state, when the flow control valve 5 is appropriately operated in a direction to open from the neutral position, the pressure oil
And the pump pressure Pd tends to decrease, so that the pump regulator 9 operates as described above, and the pump flow rate increases. The inlet pressure Pz of the flow control valve 5 and the load pressure PLS of the actuator 2 are led to the control chambers 29 and 30 of the pressure compensating valve 8 via pipes 11 and 12, respectively.
The pump pressure Pd is guided through the passage 34 to the valve spool
At 23, the inlet pressure Pz of the flow control valve 5 guided to the control chamber 29 acts in the valve closing direction, and the load pressure PLS guided to the control chamber 30 acts in the valve opening direction, and is guided to the cylinder chamber 32. Pump pressure
Pd acts in the valve opening direction.

Here, if the target value of the LS differential pressure Pd−PLS determined by the spring 14c of the pump regulator 9 is Px, and the initial target value of the compensation differential pressure of the pressure compensating valve 8 described later is Po, Px> Po is set. I have. Therefore, the valve spool 23 does not move until the LS differential pressure Pd-PLS reaches the initial target value Po of the compensation differential pressure, the pressure compensating valve 23 remains fully open, and the initial target value Po of the compensation differential pressure is reduced. After passing, the spring constant of the spring 31 and the inlet pressure Pz,
The valve spool 23 moves in the valve closing direction according to the magnitude of the load pressure PLS and the pump pressure Pd to narrow the notch 26, whereby the differential pressure across the flow control valve 5 is maintained at a predetermined value.

Next, control characteristics of the pressure compensating valve 8 of the present embodiment will be described in detail. In the operation of the pressure compensating valve 8 described above, when pressure oil flows through the notch 26 of the valve spool 23, a force acting on the valve spool 23 in the valve closing direction, that is, a flow force 36 is generated by the flow. Let this flow force 36 be f,
Assuming that the force of the spring 31 is F, the balance of the force acting on the valve spool 23 is as follows: Pz · Az + f = Pd · Ad + F + PLS · ALS (1) Pressure Pz−
Since PLS is Az = ALS + Ad, Pz−PLS = (F / Az) − (f / Az) + (Ad / Az) (Pd−PLS) (2) That is, the flow control valve 5 controlled by the pressure compensating valve 8
Is affected by the three terms on the right side of the above equation (2). Therefore, each of these three terms will be discussed below.

First, the first term on the right side of the above equation (2) is a term of pressure compensation by the spring 31, and the control characteristic by the spring 31 alone is the second term.
As shown in the figure. In the figure, the dashed line A indicates the control characteristic, and Po1 is the target value of the compensation differential pressure provided by the spring 31. That is, the LS differential pressure Pd−PLS is equal to the target value Po of the compensation differential pressure.
Until the pressure reaches 1, the valve spool 23 does not move, and the pressure compensating valve 23 remains fully open. Therefore, the differential pressure Pz-PLS before and after the flow control valve 5 changes the same as the LS differential pressure Pd-PLS. When the LS differential pressure exceeds the target value Po1 of the compensation differential pressure, the valve spool 23 moves in the valve closing direction, the notch 26 is narrowed, and the differential pressure Pz-PLS before and after the flow control valve 5 is reduced to the target value Po1 of the compensation differential pressure. To hold.

On the other hand, the second term on the right side of the above equation (2) is the flow force f with respect to the target value Po1 of the compensation differential pressure given by the spring 31.
And the flow force f is the target value Po1
Acts to reduce This flow force f is a function of the flow rate and flow rate of the pressure oil flowing through the notch 26, and the flow rate is a function of the differential pressure across the notch 26 if the flow rate is constant. Therefore, when the differential pressure Pz-PLS before and after the flow control valve 5 changes as shown in FIG. 2 with respect to the increase in the LS differential pressure Pd-PLS, the LS differential pressure is maintained until the LS differential pressure reaches Po1. As the flow rate and the flow velocity increase with the increase of the flow force f, and the LS differential pressure exceeds Po1, the differential pressure across the notch 26 increases with the increase of the LS differential pressure, The flow force f increases in accordance with the increase in the differential pressure. Eventually, the flow force f continuously increases with an increase in the LS differential pressure, and the decrease in the target value of the compensation differential pressure due to the increase in the flow force becomes as shown by a broken line B in FIG. The slope of dashed line B corresponds to -1 / Az.

The third term on the right side of the above equation (2) is a differential pressure between the pump pressure Pd and the load pressure PLS due to the provision of the cylinder chamber 32,
That is, it is a term of the influence of the LS differential pressure, and since Ad / Az = constant, L
The S differential pressure acts to increase the target value Po1 of the compensation differential pressure. The increase in the target value of the compensation differential pressure due to the LS differential pressure is as shown by a two-dot chain line C in FIG. The gradient of the two-dot chain line C is Ad
Corresponds to / Az. In the present invention, this gradient is set to be larger than the absolute value of the gradient shown by the broken line B in FIG.

The control characteristics of the above equation (2) can be obtained by synthesizing the characteristics of FIGS. 2 to 4 described above. FIG. 5 shows the synthesized characteristics by a solid line D.

As can be seen from FIG. 5, in the pressure compensating valve 8 of the present embodiment, the differential pressure Pz before and after the flow control valve 5 with respect to the LS differential pressure Pd-PLS.
The characteristic of -PLS is that when the differential pressure Pd-PLS exceeds the initial target value Po of the compensation differential pressure, the differential pressure Pd-PLS increases as the differential pressure Pd-PLS increases.
The characteristic has a positive slope in which Pz-PLS increases.

The flow rate Q flowing through the flow control valve 5 is
Let A be the opening area and K be the proportionality constant of Is generally represented by Accordingly, the characteristic of the passing flow rate Q of the flow control valve 5 with respect to the LS differential pressure corresponding to the solid line D in FIG. 5 is as shown by the solid line E in FIG. That is, LS differential pressure Pd-PLS
The characteristic of the passing flow rate Q with respect to is that the flow rate Q increases as the differential pressure Pd-PLS increases when the LS differential pressure Pd-PLS exceeds the initial target value Po of the compensation differential pressure. I have.

For comparison, the LS differential pressure of the conventional pressure compensating valve without the cylinder chamber 32 and the piston rod 33, which is a feature of this embodiment,
The characteristic of the differential pressure Pz-PLS before and after the flow control valve with respect to Pd-PLS is shown by a solid line F in FIG. 7, and the characteristic of the corresponding flow rate Q of the flow control valve corresponding to the LS differential pressure is shown by a solid line in FIG. Shown by G. As can be seen from FIG. 7, in the conventional pressure compensating valve,
When the LS differential pressure exceeds the initial target value Po2 of the compensation differential pressure, the characteristic has a negative gradient in which the differential pressure Pz-PLS decreases as the differential pressure Pd-PLS increases due to the influence of the flow force. Similarly, as shown in FIG. 8, when the LS differential pressure exceeds the initial target value Po of the compensating differential pressure, the flow rate Q decreases as the differential pressure Pd-PLS increases. .

Next, effects of the pressure compensating valve according to the present embodiment based on the above-described control characteristics will be described.

First, a hydraulic drive device using a conventional pressure compensating valve having the control characteristics shown in FIGS. 7 and 8 will be described.

When the LS differential pressure is maintained at the target value Px by the pump regulator 9, the valve spool of the pressure compensating valve moves in the valve closing direction and is in the throttle state. At this time, the characteristic line G in FIG.
The upper position is indicated by 40 and the corresponding flow rate is indicated by Q1. In this state, if the pump flow rate decreases for some reason and the flow rate passing through the flow control valve 5 decreases correspondingly, the relationship between the LS differential pressure Pd-PLS and the flow rate Q has a negative gradient. Because of the characteristic, the position 40 on the characteristic line G moves in the direction indicated by the arrow 41, and the pressure compensating valve operates to increase the LS differential pressure. On the other hand, when the LS differential pressure increases, the pump regulator 9 sets the LS differential pressure to the target value Px in response to the increase in the LS differential pressure.
To reduce the pump flow rate to maintain
When the pump flow rate decreases, the flow rate Q passing through the flow control valve 5 further decreases, the position on the characteristic line G further moves in the direction of the arrow 41, and the pressure compensating valve operates to further increase the LS differential pressure. . As a result of repeating the above, the pump flow rate is unilaterally controlled to decrease, and the pump regulator 9 is controlled.
Cannot be controlled, and the actuator cannot be driven at a desired speed.

Conversely, when the pump flow rate increases and the flow rate Q passing through the flow control valve 5 increases, the position 40 on the characteristic line G moves in the direction of the arrow 42, and the pressure compensating valve reduces the LS differential pressure. The pump regulator 9 operates so as to increase the pump flow rate in response to the decrease in the LS differential pressure, and the pressure compensating valve reduces the LS differential pressure with the increase in the flow rate Q passing through the flow control valve. It operates to further decrease. As a result of the repetition of the above, the pump flow rate is unilaterally controlled in the increasing direction, the pump regulator 9 becomes similarly uncontrollable, and the actuator cannot be driven properly.

On the other hand, in this embodiment, if the position on the characteristic line E shown in FIG. 6 when the LS differential pressure is held at the target value Px by the pump regulator 9 is 43, in this state, If the pump flow rate decreases and the flow rate through the flow control valve 5 decreases correspondingly,
Since the relationship between the LS differential pressure Pd-PLS and the flow rate Q is a characteristic having a positive gradient, the position 43 on the characteristic line E moves in the direction indicated by the arrow 44, and the pressure compensating valve 8 reduces the LS differential pressure. Operates to decrease. On the other hand, when the LS differential pressure decreases, the pump regulator 9 operates to increase the pump flow rate in order to maintain the LS differential pressure at the target value Px in response to the decrease in the LS differential pressure. When the pump flow rate increases, the flow rate Q passing through the flow control valve 5 also increases, the position on the characteristic line E moves toward the initial position 43 as indicated by the arrow 45, and the pressure compensating valve 8 sets the LS differential pressure Pd Act to return PLS to the target value Px. As a result, the LS differential pressure is again held at the target value Px.

Conversely, when the pump flow rate increases and the flow rate Q passing through the flow control valve 5 increases, the position 43 on the characteristic line E is indicated by an arrow 46.
, The pressure compensating valve 8 operates to increase the LS differential pressure, and the pump regulator 9 operates to decrease the pump flow in response to the increase in the LS differential pressure. As the passing flow rate Q decreases, the position on the characteristic line E is indicated by an arrow.
Moving in the direction of 47, the pressure compensating valve 8 operates to increase the LS differential pressure. Thereby, similarly, the LS differential pressure is again held at the target value Px.

As described above, in the present embodiment, the LS differential pressure Pd-PLS is controlled so as to return to the target differential pressure Px without being affected by the flow force.
Stable control is possible. Therefore, the flow control valve 5
Thus, the flow rate Q corresponding to the opening amount of the flow control valve 5 can be supplied to the actuator 2 via the actuator, and the drive speed of the actuator 2 can be controlled stably without being affected by the fluctuation of the pump flow rate.

Further, since no damping means such as a throttle is provided, good responsiveness of the pressure compensating valve 8 can be ensured. Further, the cylinder chamber 32 and the passage 34 are formed in the valve spool 23, and the piston rod 33 is formed in the cylinder chamber 32. Is easy to manufacture because it has a simple structure in which only a.

Second Embodiment A second embodiment of the present invention will be described with reference to FIG. In this embodiment, the means for giving the target value of the compensation differential pressure is replaced by a spring, and is constituted by hydraulic means.

In FIG. 9, the pressure compensating valve 8A of this embodiment has an input port 21a, an outlet port 21b, and two control ports 21c and 21d.
In addition, the valve housing 21A has two control ports 21e and 21f, in which a spool bore 22A, an annular inlet recess 24 and an outlet recess 25, and four control chambers 29A, 30A, 50 and 51 are formed. A valve spool 23A having a plurality of notches 26 is slidably inserted in the spool bore 21A in the axial direction.

Small-diameter portions 52, 53 are formed at both end portions of the valve spool 23A, respectively.
7A and 54 are formed, and pressure receiving portions 55 and 28A are formed on end faces of the small diameter portions 52 and 53. The pressure receiving parts 27A, 28A are in control rooms 29A, 3 respectively.
0A, and these control rooms 29A, 30A have control ports 21c, 2c.
The inlet pressure Px of the flow control valve 4 and the load pressure PLS of the actuator 2 are led through 1d, respectively. Pressure receiving part 5
4 and 55 are located in the control rooms 50 and 51, respectively, and the control room 50 is connected to the hydraulic pressure source 56 through the control port 21e.
Is connected to an electromagnetic proportional valve 58 connected to a hydraulic pressure source 57 via a control port 21f.

The hydraulic pressure sources 56 and 57 each generate a constant pilot pressure Pi. In addition, the electromagnetic proportional valve 58 reduces a constant pilot pressure from the hydraulic pressure source 57 according to the electric signal, and generates a control pressure Pc according to the electric signal. The control force generated in the control chamber 50 by the pilot pressure Pi from the hydraulic pressure source 56 is applied to the valve spool.
23A is urged in the valve opening direction, and the control pressure P from the electromagnetic proportional valve 58
The control force generated in the control chamber 51 by c urges the valve spool 23A in the valve closing direction. The pressure receiving areas of the pressure receiving sections 54 and 55 are equal as described later, and the pilot pressure Pi and the control pressure
Pc is set so that the former control force is greater than the latter control force, whereby the difference between the two control forces urges the valve spool 23A in the valve opening direction, and the difference between the control force and the spring 31 of the first embodiment. Similarly, a target value of the compensation differential pressure is given. Also, by controlling the electromagnetic proportional valve 58 to adjust the control pressure Pc, the difference between the two control forces can be controlled, and the target value of the compensation differential pressure can be freely changed.

The control of the electromagnetic proportional valve is performed, for example, in EP, A1, 326, 15
0 (corresponding to Japanese Patent Application Laid-Open No. 1-312202), whereby a hydraulic drive device for driving a plurality of actuators can set a target value of compensation differential pressure of a plurality of pressure compensating valves when a hydraulic pump is saturated. Appropriate flow rate control such as split flow control for surely supplying pressure oil to each actuator can be performed by appropriately changing each of them.

The valve spool 23A has a cylinder chamber 3 as in the first embodiment.
2. A passage 34 and a pressure receiving portion 35 are formed, and a piston rod 33 is inserted into the cylinder chamber 32.

The pressure receiving area of the pressure receiving section 27A is Az, and the pressure receiving area of the pressure receiving section 28A is A.
Assuming that LS, the pressure receiving area of the pressure receiving section 35 is Ad, the pressure receiving area of the pressure receiving section 54 is Ai, and the pressure receiving area of the pressure receiving section 55 is Ac, Az + Ac = ALS
+ Ad + Ai and Az = ALS + Ad (accordingly, Ac = Ai).

In the above configuration, the control chambers 29A, 30A and the pressure receiving sections 27A, 2
8A is a differential pressure Pz-PLS between the front and rear of the flow control valve 5 on the valve spool 23A.
To provide a first control means for applying a first control force based on the control signal, and for urging the valve spool 23A in the valve closing direction.
The pressure receiving portions 54 and 55 provide a second control unit that applies a second control force based on the pilot pressure Pi and the control pressure Pc to the valve spool 23A and urges the valve spool 23A in the valve opening direction. . The control chamber 29A and the pressure receiving portion 27A, the cylinder chamber 32 and the pressure receiving portion 35 are provided with a differential pressure between the pump pressure and the load pressure controlled by the pump regulator 9 LS, that is, the LS differential pressure Pd.
-A third control means for applying a third control force, which increases in accordance with an increase in PLS, to the valve spool 23A and biasing the valve spool 23A in the valve opening direction is provided.

In the present embodiment configured as above, the balance of the forces acting on the valve spool 23A in consideration of the flow force f is Pc · Ac + Pz · Az + f = Pd · Ad + Pi · Ai + PLS · ALS (4) From equation 4), the differential pressure Pz-
Since PLS is Az + Ac = Ad + ALS + Ai and Ac = Ai, Pz−PLS = {Ai (Pi−Pc) / Az} − (f / Az) + (Ad / Az) (Pd−PLS) 5) In the equation (5), the first term on the right side is the control pressure.
This value is determined according to Pi, and corresponds to the first term on the right side of Expression (2). The second and third terms on the right side are the same as those in equation (2).

Therefore, also in this embodiment, the characteristic of the differential pressure Pz-PLS before and after the flow control valve 5 controlled by the pressure compensating valve 8A with respect to the LS differential pressure Pd-PLS is as shown by a solid line D in FIG.
Similarly, the characteristic of the passing flow rate Q of the flow control valve 6 with respect to the LS differential pressure is as shown by a solid line E in FIG. That is, LS differential pressure Pd
The characteristic of the differential pressure Pz-PLS before and after the flow control valve 5 with respect to PLS is that when the LS differential pressure Pd-PLS exceeds the initial target value Po of the compensation differential pressure, the differential pressure Pz-PLS increases as the differential pressure Pd-PLS increases. The characteristic has a positive slope in which PLS increases. Also, the LS differential pressure Pd-P
The characteristic of the flow rate Q of the flow control valve 5 with respect to LS is the LS differential pressure.
When Pd-PLS exceeds the initial target value Po of the compensation differential pressure, the differential pressure Pd
-A characteristic having a positive slope in which the flow rate Q increases as PLS increases.

Therefore, in this embodiment, the same effect as that of the first embodiment can be obtained. Further, according to the present embodiment, by changing the control pressure Pc, the target value Po1 of the compensation differential pressure shown in FIG. 2 can be appropriately changed, whereby the desired flow rate control such as the above-described split flow control at the time of pump saturation is achieved. It can be performed.

Third Embodiment A third embodiment of the present invention will be described with reference to FIG. In this embodiment, means for applying a control force that increases in accordance with an increase in the LS differential pressure is provided outside the valve spool, not inside the valve spool as in the pressure receiving section in the cylinder chamber.

In FIG. 10, the pressure compensating valve 8B of this embodiment has a valve housing 21B having an inlet port 21a, an outlet port 21b, and two control ports 21c and 21d, as in the first embodiment,
A spool bore 22B, an annular inlet recess 24 and an outlet recess 25, and two control chambers 29B and 30B are formed in the valve housing 21B. A valve spool 23B having a plurality of notches 26 is slidably inserted in the spool bore 21B in the axial direction.

One end of the valve spool 23B has a large-diameter portion 60, and a pressure-receiving portion 27B is formed on an end surface of the large-diameter portion 60, and a pressure-receiving portion 28B is formed on an end surface of the other end. Pressure receiving part 27B, 28B
Are located in the control rooms 29B and 30B, respectively,
Also has a larger diameter than the control room 29A. The inlet pressure Pz of the flow control valve 4 and the load pressure PLS of the actuator 2 are led to the control chambers 29B and 30B via control ports 21c and 21d, respectively. Further, a spring 31 is disposed in the control room 30B.

The inlet recess 2 is formed in the shoulder 61 of the large diameter portion 60 on the opposite side of the pressure receiving portion 27B.
An annular end face facing 4 is formed, and this end face has an inlet recess 24.
A pressure receiving portion 62 is formed to receive the pump pressure of the above and urge the valve spool in the valve opening direction.

The pressure receiving area of the pressure receiving section 27B is Az, and the pressure receiving area of the pressure receiving section 28B is A
Assuming that LS and the pressure receiving area of the pressure receiving section 62 are Ad, the relationship is Az = ALS + Ad.

In the above configuration, the control chambers 29B, 30B and the pressure receiving sections 27B, 2
The functions of 8B and the spring 31 are the same as in the first embodiment. Further, the control chamber 29B and the pressure receiving section 27B and the pressure receiving section 62 have a third pressure that increases in accordance with an increase in the differential pressure between the pump pressure and the load pressure controlled by the pump regulator 9 in LS, that is, the LS differential pressure Pd-PLS. The control force is applied to the valve spool 23B, and the valve spool 23B
The third control means for urging the valve in the valve opening direction is provided.

In the present embodiment configured as above, the balance of the force acting on the valve spool 23B in consideration of the flow force f is Pz · Az + f = Pd.Ad + F + PLS · ALS, which is the same as the above equation (1). Differential pressure Pz-PLS before and after control valve 5
Since Az = ALS + Ad, Pz-PLS = (F / Az)-(f / Az) + (Ad / Az) (Pd-PLS), which is the same as the equation (2).

Therefore, also in this embodiment, the characteristic of the differential pressure Pz-PLS before and after the flow control valve 5 controlled by the pressure compensating valve 8B with respect to the LS differential pressure Pd-PLS is positive as shown by the solid line D in FIG. A characteristic having a gradient, and a characteristic of the passing flow rate Q of the flow control valve 6 with respect to the LS differential pressure also has a characteristic having a positive gradient as shown by a solid line E in FIG. Therefore, according to this embodiment, the same effect as that of the first embodiment can be obtained.

Industrial Applicability According to the present invention, the pressure difference between the pump pressure and the load pressure (LS
Since the characteristic of the flow rate of the flow control valve with respect to the differential pressure) has a positive slope, that is, the characteristic that the LS differential pressure increases and decreases in response to the increase and decrease of the flow rate, the flow rate makes the pump flow rate uncontrollable. Therefore, stable flow control can be performed. In addition, since good responsiveness can be ensured and the configuration is relatively simple, manufacturing is easy.

────────────────────────────────────────────────── ─── Continuation of the front page (56) References JP-A-58-39807 (JP, A) JP-A-61-226495 (JP, A) JP-A-1-321201 (JP, A) (58) Field (Int.Cl. ⁶ , DB name) F15B 11/00 F15B 11/05 E02F 9/22

Claims (13)

(57) [Claims] 1. A hydraulic pump (1), an actuator (2) driven by pressure oil discharged from the hydraulic pump, and a flow control valve (5) disposed between the hydraulic pump and the actuator. A pressure compensating valve (8, 8A, 8B) for controlling the pressure difference (Pz-PLS) of the flow control valve, and the flow rate discharged from the hydraulic pump is determined by the difference between the pump pressure and the load pressure of the actuator. And a pump flow control means (9) for controlling the pressure compensating valve according to the pressure (Pd-PLS). The pressure compensating valve comprises a valve element (23, 23A, 23B), Providing a first control force based on pressure;
The first control means (29, 30; 29A,
30A; 29B, 30B) and a second control means (31; 5) for applying a predetermined second control force to the valve body and urging the same in the valve opening direction.
0,51), the pressure compensating valves (8, 8A, 8B) are based on a differential pressure (Pd-PLS) between the pump pressure and the load pressure of the actuator. A third control force that continuously increases as the differential pressure (Pd-PLS) increases, is applied to the valve body (23, 23A, 23B), and a third control force is urged in the valve opening direction. Control means (32, 35;
62) A hydraulic drive system for civil engineering and construction machinery, further comprising: 2. The hydraulic drive device for civil engineering and construction machinery according to claim 1, wherein said third control means is adapted to open said valve element (23, 23A, 23B) by being guided by said pump pressure (Pd). A hydraulic drive device for a civil engineering / construction machine, comprising: a first pressure receiving means (35; 62) biasing in a valve direction. 3. The hydraulic drive system for civil engineering and construction machinery according to claim 2, wherein said first pressure receiving means (35) is formed inside said valve element (23, 23A). Hydraulic drive for civil engineering and construction machinery. 4. The hydraulic drive system for a civil engineering and construction machine according to claim 2, wherein said first pressure receiving means (62) is formed outside said valve element (23B).・ Hydraulic drive for construction machinery. 5. The hydraulic drive system for civil engineering and construction equipment according to claim 1, wherein said first control means includes a first control chamber (1) into which an inlet pressure (Pz) of said flow control valve (5) is introduced. 29; 2
9A; 29B) and a first pressure receiving portion (2) disposed in the first control chamber and for urging the valve element (23, 23A; 23B) in a valve closing direction.
7; 27A; 27B), a second control chamber (30; 30A; 30B) into which the load pressure (PLS) is led, and the second control chamber is disposed in the second control chamber. Second pressure receiving part (28; 2
8A; 28B), and the third control means includes a third pressure receiving portion (35; 62) to which the pump pressure (Pd) acts and urges the valve body in the valve opening direction. Civil engineering wherein the sum of the pressure receiving area (ALS) of the second pressure receiving section and the pressure receiving area (Ad) of the third pressure receiving section is substantially equal to the pressure receiving area (Az) of the first pressure receiving section.・ Hydraulic drive for construction machinery. 6. A hydraulic drive system for a civil engineering / construction machine according to claim 5, wherein said third control means includes a valve body (23, 2).
3A), a cylinder chamber (32) formed in the axial direction, one end of which is closed and the other end of which is open to the second pressure receiving portion (28; 28A), and is slidably inserted into the cylinder chamber. A piston rod protruding out of the valve body from the other end;
And a passage (34) formed in the valve body and guiding the pump pressure (Pd) to the cylinder chamber. The third pressure receiving portion (35) is formed at a closed end of the cylinder chamber. A hydraulic drive device for civil engineering and construction machinery. 7. A hydraulic drive system for civil engineering and construction machinery according to claim 5, wherein said valve element (23B) has a large diameter portion (6) at one end thereof.
0), and the first pressure-receiving portion (27B) is formed on the end face of the large-diameter portion, and the third control means controls the large-diameter portion on the side opposite to the first pressure-receiving portion (61B). ), Wherein the third pressure receiving portion (62) is formed on the annular end surface. 8. The hydraulic drive device for a civil engineering / construction machine according to claim 1, wherein said second control means is a spring (31). 9. The hydraulic drive system for civil engineering and construction machinery according to claim 1, wherein said second control means is a hydraulic means (50, 51) for generating said second control force by hydraulic pressure. Hydraulic drive for civil engineering and construction machinery. 10. A hydraulic drive system for a civil engineering / construction machine according to claim 9, wherein said hydraulic means is a first hydraulic pressure generating means (56) for generating a constant hydraulic pressure, and said constant hydraulic pressure is guided, A second pressure receiving means (50) for urging the valve element (23A) in a valve opening direction, a second oil pressure generating means (57, 58) for generating a variable oil pressure, and And a third pressure receiving means (51) for guiding and urging the valve body in a valve closing direction. 11. A pressure compensating valve (8; 8A; 8B) for controlling a pressure difference between before and after a flow control valve (5) disposed between a hydraulic pump (1) and an actuator (2), comprising: A valve housing (21; 21A; 21B) having a bore (22; 22A, 22B), an inlet recess (24) connected to the hydraulic pump, and an outlet recess (25) connected to the flow control valve; A valve spool (23; 24A; 23B) slidably disposed in the spool bore and controlling communication between the inlet recess and the outlet recess;
A first control chamber (29; 29A; 29B) formed in the valve housing to guide the inlet pressure of the flow control valve; and a first control chamber disposed in the first control chamber, wherein the valve spool is moved in a valve closing direction. A first pressure receiving portion (27; 27A; 27B) for energizing, a second control chamber (30; 30A; 30B) formed in the valve spool, and to which a load pressure of the actuator is led; A second pressure receiving portion (28; 28A; 28B) for urging the valve spool in the valve opening direction, and urging the valve spool in a valve opening direction with a predetermined control force to compensate A pressure compensating valve having means (31; 50, 51) for setting a target value of the differential pressure, wherein the pressure in the inlet recess (24) is led to open the valve spool (23; 23A; 23B). Pressure receiving part (3
5; 62), the pressure receiving area (ALS) of the second pressure receiving section (28; 28A; 28B) and the pressure receiving area (Ad) of the third pressure receiving section (35; 62). The pressure compensation area is substantially equal to the pressure receiving area (Az) of the first pressure receiving portion (27; 27A; 27B). 12. The pressure compensating valve according to claim 11, wherein said control means is formed in the valve spool (23, 23A) in an axial direction, one end of which is closed, and the other end of which is the second pressure receiving valve. A cylinder chamber opening to the portion (28; 28A); a piston rod (33) slidably inserted into the cylinder chamber (32) and projecting out of the valve spool from the other end; A passage (34) formed to guide the pressure of the inlet recess into the cylinder chamber, and the third pressure receiving portion (35)
The pressure compensating valve is formed at a closed end of the cylinder chamber. 13. The pressure compensating valve according to claim 11, wherein said valve spool (23B) includes a large-diameter portion (60) at one end thereof, and said first pressure-receiving portion (27B) is provided at an end surface of said large-diameter portion. )
Is formed on the opposite side of the large-diameter portion from the first pressure receiving portion (61), and includes an annular end face facing the inlet recess, and the third pressure receiving portion (62 ) Is a pressure compensating valve formed on the annular end face.