JP2002364549A - Capacity control device for hydraulic pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a capacity control device for a hydraulic pump having simple and small construction and allowing the effective use of horse power output by a power source. SOLUTION: Two flow control mechanisms 11, 12 and a common horse power control mechanism 13 located therebetween are provided for a regulator 7 and feedback sleeves 11d, 12d of the flow control mechanisms 11, 12 are linked to servo pistons 3a, 4a in differential cylinders 3, 4 via feedback links 17, 18 and linked to a feedback sleeve 13d of the horse power control mechanism 13 via an intermediate link 19. Two flow control mechanisms 11, 12 are each arranged at an equal space from the horse power control mechanism 13 and a link mechanism consisting of links 17, 18, 19 is constructed so that the moving quantity of the feedback sleeve 13d of the horse power control mechanism 13 is set to be an average value for the moving quantity of the feedback sleeves 11d, 12d of the flow control mechanisms 11, 12.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic pump displacement control device, and more particularly to a hydraulic pump displacement control device for controlling a discharge flow rate of at least two identically configured variable displacement hydraulic pumps driven by one power source. .

[0002]

2. Description of the Related Art A hydraulic drive device of a hydraulic construction machine such as a hydraulic excavator is provided with a variable displacement type hydraulic pump as a hydraulic source for supplying pressure oil to an actuator. Usually, two or more variable displacement hydraulic pumps having the same configuration are provided, and the two or more hydraulic pumps are driven by one power source (usually an engine). These hydraulic pumps need to discharge a flow rate corresponding to the required flow rate. In addition, when the discharge pressure of the hydraulic pump increases (actuator load increases), when the actuator load is excessively large, the discharge flow rate is reduced to reduce engine stall. A hydraulic pump displacement control device is provided for this purpose.

[0003] Conventionally, as a hydraulic pump displacement control device,
The one shown in FIG. 25 has been put to practical use. FIG. 26 is a hydraulic circuit diagram thereof.

The hydraulic pump displacement control devices shown in FIGS. 25 and 26 control the discharge flow rates of two hydraulic pumps 101 and 102, and the hydraulic pumps 101 and 102 have the same structure. Further, the hydraulic pump displacement control device is provided with operating cylinders 103 and 104 having the same structure and a regulator 10 for each of the hydraulic pumps 101 and 102.
5 and 106, the working cylinders 103 and 104 have servo pistons 103a and 104a, respectively, and the servo pistons 103a and 104a are hydraulic pumps 101 and 10 respectively.
2 are linked to the displacement variable mechanisms 101a and 102b. The regulators 105 and 106 respectively have flow control mechanisms 111 and 112 and a horsepower control mechanism 113,
114, and flow control command pressures Pq1 and Pq2 corresponding to the required flow rates of the hydraulic pumps 101 and 102 are input to the pressure receiving units 117 and 118 of the flow control mechanisms 111 and 112, respectively.
Pressure receiving units 115a, 115 of horsepower control mechanisms 113, 114
b; The discharge pressures Pd1 and Pd2 of the hydraulic pumps 1 and 2 are input to 116a and 116b as horsepower control command pressures.

[0005] The differential cylinders 103 and 104 have small-diameter pressure chambers 103b and 104b and large-diameter pressure chambers 103c and 104, respectively.
c, and the pressure Pg of the pilot hydraulic pressure source is constantly guided to the small-diameter pressure chambers 103b and 104b.
The tank pressure Dr and the pressure Pg of the pilot hydraulic pressure source are selectively guided to c and 104c. Large-diameter pressure chamber 103c, 1
When the tank pressure Dr is guided to the pressure chamber 04c, the servo pistons 103a and 104a move to the left in the figure to increase the discharge flow rates of the hydraulic pumps 101 and 102, and the pilot pressure sources are supplied to the large-diameter pressure chambers 103c and 104c. When the pressure Pg is guided, the servo pistons 103a and 104a move rightward in the figure to decrease the discharge flow rates of the hydraulic pumps 101 and 102.

The flow control mechanisms 111 and 112 and the horsepower control mechanisms 113 and 114 of the regulators 105 and 106 are connected to the large-diameter pressure chambers 103 c and 103 of the dynamic cylinders 103 and 104.
4c to the tank pressure Dr or the pilot pressure P
An operation is performed to control the switching of whether or not to guide g, whereby the discharge flow rate of the hydraulic pumps 101 and 102 is controlled. The flow control mechanism 111 of the regulator 105 operates such that the discharge flow rate of the hydraulic pump 1 increases as the flow control command pressure Pq1 increases, and the discharge flow rate of the hydraulic pump 1 decreases as the flow control command pressure Pq1 decreases. The cylinder 103 is controlled, and the horsepower control mechanism 113 controls the operating cylinder 1 to decrease the discharge flow rate of the hydraulic pump 1 as the discharge pressures Pd1 and Pd2 of the hydraulic pumps 1 and 2 increase.
03 is controlled. Flow control mechanism 1 of regulator 106
12 and the horsepower control mechanism 114 are the same. The operating principle of each regulator is described in Japanese Patent Application Laid-Open No. H11-14846.
No. 3 is detailed.

[0007]

As described above, the conventional hydraulic pump displacement control apparatus is capable of controlling the flow control command pressures Pq1 and Pq1.
q2, the discharge flow rate of the hydraulic pumps 101 and 102 is increased or decreased to obtain a discharge flow rate corresponding to the required flow rate, and when the horsepower control command pressures Pd1 and Pd2 increase, the hydraulic pump 10
The discharge flow rate of the discharge pumps 1, 102 is reduced to prevent the engine from being stalled due to an increase in the discharge pressure of the hydraulic pump (an increase in the actuator load).

[0008] However, the conventional hydraulic pump displacement control device has two hydraulic pumps 101 and 102 as described above.
Are provided with respective regulators 105 and 106, and each of the regulators 105 and 106 is provided with a flow control mechanism 11
1, 112 and horsepower control mechanisms 113, 114. For this reason, there is a problem that the structure becomes complicated and the size becomes large.

The horsepower control mechanisms 113 and 114 include pressure receiving portions 115a and 115b; 116a and 116, respectively.
b: Horsepower control command pressure (discharge pressure of hydraulic pumps 1, 2)
Pd1 and Pd2 are input, and hydraulic pumps 1 and 2, respectively.
Control is performed so that the discharge flow rate of the hydraulic pumps 1 and 2 is reduced as the discharge pressures Pd1 and Pd2 of the second pump 2 increase (total horsepower control). That is, the horsepower control mechanisms 113 and 114 perform the same control on the hydraulic pumps 101 and 102. For this reason, at the time of horsepower control, the hydraulic pumps 101 and 102 can always discharge only the same flow rate, and there is a problem that not all of the engine horsepower can be used effectively.

FIG. 27 shows a horsepower control characteristic diagram of the hydraulic pumps 101 and 102 by the horsepower control mechanisms 113 and 114.
FIG. 27A shows the hydraulic pump 101, and FIG. 27B shows the hydraulic pump 102. The horizontal axis is the hydraulic pump 101,
The average value (average discharge pressure) Pdm (Pdm = Pd1 + Pd2 / 2) of the discharge pressures Pd1 and Pd2 of 102 is shown, and the vertical axis is the discharge flow rates Q1 and Q2 of the hydraulic pumps 101 and 102.
H1 and H2 denote PQ lines for horsepower control, and horsepower control springs 113a and 113a of horsepower control mechanisms 113 and 114, respectively.
13b; set by 114a and 114b (see FIG. 25). Qmax is the maximum discharge flow rate determined by the mechanical limitations of the hydraulic pumps 101 and 102.

For example, when the average value Pdm of the discharge pressures of the hydraulic pumps 101 and 102 is Pc, the hydraulic pump 10
The maximum discharge flow rate of 1,102 is Qc on the PQ lines H1, H2. Therefore, suppose that the required flow rate of the hydraulic pump 101 is Qp1 larger than Qc,
When the required flow rate of 02 is Qp2 smaller than Qc, the discharge flow rate of the hydraulic pump 101 is limited to Qc equal to or less than the required flow rate Qp1, and the discharge flow rate of the hydraulic pump 102 becomes Q2p (<Qc) as the required flow rate. . In other words, for the purpose of horsepower control, the hydraulic pumps 101 and 102 can discharge a flow up to 2Qc, but discharge only a flow of Qc + Q2p, and cannot use the engine output horsepower of Qc-Q2p.

A first object of the present invention is to provide a hydraulic pump displacement control device which has a simple structure and can be reduced in size.

A second object of the present invention is to provide a hydraulic pump displacement control device capable of effectively utilizing the output horsepower of a power source.

[0014]

(1) In order to achieve the first object, the present invention provides a displacement capacity variable mechanism of at least two variable displacement hydraulic pumps driven by one power source. At least two differential cylinders that operate, and a regulator that controls the supply and discharge of pressure oil between these differential cylinders, a hydraulic source, and a tank, wherein the regulator is operated by a flow control command pressure, And a horsepower control means that operates according to the horsepower control command pressure. The flow rate control means and the horsepower control means control the servo pistons of the two differential cylinders and control the displacement mechanism of the two hydraulic pumps. In the hydraulic pump displacement control device, the flow control means includes:
It has at least two flow control mechanisms for inputting respective flow control command pressures of the two hydraulic pumps and individually operating the two differential cylinders, and the horsepower control means includes:
It is assumed that a common horsepower control mechanism for inputting the horsepower control command pressure of each of the at least two hydraulic pumps and operating the two differential cylinders is provided.

By providing a common horsepower control mechanism as the horsepower control means, the structure is simplified and the size can be reduced.

(2) In order to achieve the second object, the present invention provides the hydraulic pump displacement control device of (1), wherein the at least two flow control mechanisms and the common horsepower control mechanism are , Each feedback sleeve and a link mechanism for linking the feedback sleeves to each other and to the servo piston of the differential cylinder.

By providing the link mechanism as described above and linking the respective feedback sleeves to each other and to the servo piston of the differential cylinder, the horsepower control mechanism is considered when considering the horsepower control of each of at least two hydraulic pumps. The position of the feedback sleeve varies depending not only on the position of the feedback sleeve of the flow control mechanism relating to itself, but also on the position of the feedback sleeve of the flow control mechanism on the other side. A variable horsepower control characteristic that changes the horsepower control characteristic can be obtained. For this reason, when there is a margin in the pump discharge flow rate on the partner side with respect to the limit value of the horsepower control, the pump discharge flow rate of the pump itself can be increased by the flow rate corresponding to the margin, and the output horsepower of the power source is effectively used. be able to.

(3) In the above (2), preferably,
The common horsepower control mechanism is disposed between the at least two flow control mechanisms, and the link mechanism is configured to move the feedback sleeve of the common horsepower control mechanism by moving the feedback sleeve of the at least two flow control mechanisms. It is configured to be an average of the quantities.

Thus, as described in the above (2), when there is a margin in the other party's pump discharge flow rate with respect to the horsepower control limit value, the own pump discharge flow rate can be increased by the extra flow rate. And the output horsepower of the power source can be used effectively.

[0020]

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

FIG. 1 is a structural view of a hydraulic pump displacement control device according to an embodiment of the present invention, and FIG. 2 is a hydraulic circuit diagram showing the operation principle.

1 and 2, the hydraulic pump displacement control device according to the present embodiment has two differential cylinders 3,
4 and a common regulator 7, and an operating cylinder 3,
Numeral 4 is linked to variable displacement mechanisms 1a and 2a of the variable displacement hydraulic pumps 1 and 2 shown in FIG.
Drives the variable displacement mechanisms 1a and 2a by controlling the differential cylinders 3 and 4, and controls the displacement (capacity) of the hydraulic pumps 1 and 2.

In FIG. 2, hydraulic pumps 1 and 2 are provided as hydraulic sources for supplying pressure oil to an actuator (not shown), and each has the same configuration and one power source (not shown), for example, a diesel engine. Driven by The hydraulic pumps 1 and 2a are controlled by the operation cylinders 3 and 4 to control the variable volume mechanisms 1a and 2a.
2 is controlled.

Returning to FIG. 1, the differential cylinder 3 has a small-diameter pressure chamber 3b and a large-diameter pressure chamber 3c formed by a servo piston 3a.
The port 3d, which is the oil inlet / outlet of the small diameter side pressure chamber 3b, is connected to the hydraulic pressure source 5, and the port 3e, which is the oil inlet / outlet of the large diameter side pressure chamber 3c, is connected to the actuator port 7a of the regulator 7. Have been.

Similarly, the differential cylinder 4 is divided into a small-diameter side pressure chamber 4b and a large-diameter side pressure chamber 4c by a servo piston 4a. The port 4e of the side pressure chamber 4c is connected to the actuator port 7b of the regulator 7.

The regulator 7 has two hydraulic pumps 1,
Flow control command pressure Pq corresponding to each required flow rate of 2
1, Pq2, and the two flow rate control mechanisms 11, 12, which individually operate the two differential cylinders 3, 4, and the discharge pressures Pd1, Pd2 of the two hydraulic pumps 1, 2, respectively. Hereinafter, the horsepower control command pressures Pd1 and Pd
2) and a common horsepower control mechanism 13 for operating the two differential cylinders 3 and 4, and the position of the servo piston 3 a of the differential cylinder 3 is controlled by the flow control mechanism 11 and the horsepower control mechanism 13. By controlling the position (tilt angle) of the variable displacement mechanism 1a of the hydraulic pump 1, the discharge flow rate is controlled, and the flow control mechanism 12 and the horsepower control mechanism 13
By controlling the position of the servo piston 4a of the differential cylinder 4, the position (tilt angle) of the displacement volume variable mechanism 2a of the hydraulic pump 2 is controlled, and the discharge flow rate is controlled. The horsepower control mechanism 13 is disposed at the center between the two flow control mechanisms 11 and 12. Flow control command pressure Pq1, P
As q2, a pressure proportional to the required flow rate, such as a pilot pressure from the operation lever, is given.

The regulator 7 has a casing 15, and the flow control mechanisms 11, 12 and the horsepower control mechanism 13 are incorporated in the casing 15. The casing 15 includes, in addition to the actuator ports 7a and 7b, hydraulic source ports 7c and 7d connected to the hydraulic source 5,
A tank port 7e connected to the tank 6, a flow control port 7f for inputting flow control command pressures Pq1 and Pq2,
7g (see FIG. 2) and the horsepower control command pressures Pd1 and Pd2
And the horsepower control ports 7h and 7i (see FIG. 2) for inputting.

The flow control mechanism 11 includes a flow control spool 1
1a, a command pressure receiving portion 11b, and a flow control spring 11c.
And a flow control feedback sleeve 11d,
The above-mentioned flow control command pressure Pq1 is applied to the command pressure receiving portion 11b.
Is entered. The flow control spool 11a is provided with two lands 11a1 and 11a2, and the flow control feedback sleeve 11d is opened and closed by a control port 11e opened and closed by the land 11a1, an actuator port 11f which is normally open, and a land 11a2. A hydraulic pressure source port 11g is formed.

The same applies to the flow control mechanism 12. Flow control spool 12a, command pressure receiving portion 12b, flow control spring 12c, flow control feedback sleeve 12d
The flow control command pressure Pq2 is input to the command pressure receiving unit 12b. The flow control spool 12a is provided with two lands 12a1 and 12a2, and the flow control feedback sleeve 12d is provided with the lands 12a1 and 12a2.
A control port 12e that is opened and closed by a1, an actuator port 12f that is normally open, and a hydraulic power source port 12g that is opened and closed by a land 12a2 are formed.

The horsepower control mechanism 13 includes a horsepower control spool 1.
3a, command pressure receiving portions 13b1 and 13b2, horsepower control springs 13c1 and 13c2, and a horsepower control feedback sleeve 13d.
The above-mentioned horsepower control command pressures Pd1 and Pd2 are input to b2. The input of the horsepower control command pressures Pd1 and Pd2 to the command pressure receiving portions 13b1 and 13b2 is caused by the horsepower control command pressures Pd1 and Pd1 being applied to the command pressure receiving portions 13b1 and 13b2.
It is equivalent to inputting the average value of d2, that is, the average value (average discharge pressure) of the discharge pressures of the hydraulic pumps 1 and 2,
Thereby, so-called total horsepower control can be performed. The horsepower control spool 13a has two lands 13a1 and 13a2, and the horsepower control feedback sleeve 13d has
Tank port 13 opened and closed by land 13a1
e, a normally open control port 13f and a hydraulic power source port 13g opened and closed by the land 13a2 are formed.

The flow control feedback sleeve 11d,
The control ports 11e and 12e of the 12d communicate with the control port 13f of the horsepower control feedback sleeve 13d through the internal passages 20 and 21 of the casing 15, and the actuator ports 11f and 12f of the flow control feedback sleeves 11d and 12d are connected to the casing 15 respectively. Hydraulic source port 11 of flow rate control feedback sleeves 11d and 12d communicates with actuator ports 7a and 7b.
g, 12g and horsepower control feedback sleeve 13d
The hydraulic power source port 13g communicates with the hydraulic power source ports 7c and 7d of the casing 15 directly or through the internal passages 22 and 23 of the casing 15.

The feedback sleeve 11d of the flow control mechanism 11 is linked to the servo piston 3a of the differential cylinder 3 via a feedback link 17 and the feedback sleeve 12d of the flow control mechanism 12 feeds back to the servo piston 4a of the differential cylinder 4. Link 1
The feedback sleeves 11 d and 12 d of the flow control mechanisms 11 and 12 and the feedback sleeve 13 d of the horsepower control mechanism 13 are linked to each other via one intermediate link 19.

The arrangement intervals of the two flow control mechanisms 11 and 12 and the horsepower control mechanism 13 are equal, and
8 and 19, the movement amount of the feedback sleeve 13 d of the horsepower control mechanism 13 is controlled by the flow control mechanism 1.
The first and second feedback sleeves 11d and 12d are configured to have an average value of the movement amount. For example, when the servo pump 3a of the differential cylinder 3 moves when the hydraulic pump 1 is operated alone, the flow control feedback sleeve 11d moves at a fixed ratio to the movement amount of the servo piston 3a. At this time, the movement of the flow rate control feedback sleeve 11d causes the intermediate link 19 to rotate around the link connection point with the flow rate control feedback sleeve 12d, and the horsepower control feedback sleeve 13d moves by half the moving distance of the flow rate control feedback sleeve 11d. . The same applies when the hydraulic pump 2 is operated alone. When both of the hydraulic pumps 1 and 2 are operated, the flow rate control feedback sleeve 11d by the movement of the differential cylinder 3
Is the amount of movement of the flow control feedback sleeve 12d due to the movement of the differential cylinder 4, and the intermediate link 19 is the movement control feedback sleeve 11
d and 12d, the horsepower control feedback sleeve 13d moves by (x + y) / 2.

In the hydraulic pump displacement control device configured as described above, the relative positional relationship between the flow control spool 11a and the flow control feedback sleeve 11d, the flow control spool 12a and the flow control feedback sleeve 1d.
The relative position relationship between the control port 11e and the hydraulic power source port 11g of the flow control feedback sleeve 11d and the control port of the flow control feedback sleeve 12d are determined by the relative positional relationship between the control port 11e and the hydraulic power source port 11g. 12e and hydraulic source port 12g open / close state, tank port 13e of horsepower control feedback sleeve 13d and hydraulic source port 1
The open / close state of 3g is determined, and the combination thereof determines whether the servo pistons 3a and 4a move in the increasing direction or the decreasing direction of the pump discharge flow rate, or stop.

As described above, the two flow control mechanisms 1
1, 12 and the horsepower control mechanism 13 are arranged at equal intervals, and the link mechanism including the links 17, 18, 19 has a movement amount of the feedback sleeve 13d of the horsepower control mechanism 12 that is equal to the feedback sleeve 11d of the flow control mechanisms 11, 12.
12d, the actual discharge flow rates of the hydraulic pumps 1 and 2 are set to Q1, Q2,
The required flow rate related to the flow control command pressures Pq1 and Pq2 is Q
1p, Q2p, and two hydraulic pumps 1,
Assuming that the average discharge flow rate of No. 2 is Qc, it operates in the flow rate control mode and the horsepower control mode as described below according to the magnitude relation.

A: Q1p + Q2p ≦ 2Qc → flow rate control mode Q1 = Q1p Q2 = Q2p B: Q1p + Q2p> 2Qc → horsepower control mode Q1p ≧ Qc, Q2p ≧ Qc Q1 = 2Qc−Q2 Q2 = 2Qc−Q1 For example, when Q1 = Qc, = Qc If the average discharge pressures of the two hydraulic pumps 1 and 2 are constant, Q1 and Q2 are also constant. Q1p> Qc, Q2p <Qc Q1 = 2Qc−Q2 ≦ Qmax (Qmax is the upper limit in mechanism) Q2 = Q2p 2 Even if the average discharge pressure of the two hydraulic pumps 1 and 2 is constant, Q1 changes with the change of Q2. (Variable horsepower control) Q1p <Qc, Q2p> Qc Q1 = Q1p Q2 = 2Qc-Q1 ≦ Qmax Even if the average discharge pressure of the two hydraulic pumps 1 and 2 is constant, Q2 changes with the change of Q1. (Variable horsepower control) In this specification, the horsepower control of B is referred to as variable horsepower control.

Hereinafter, the operation of the hydraulic pump displacement control device will be described in detail.

First, the flow control mode and the horsepower control mode, which are basic operations, will be described.

<Flow control mode> The flow control mode is as follows.
This is the case where Q1p + Q2p ≦ 2Qc as described above. In this case, the relative positional relationship between the horsepower control spool 13a of the horsepower control mechanism 13 and the feedback sleeve 13d shown in FIG. 1 is such that the tank port 13e of the feedback sleeve 13d is opened and the hydraulic pressure source port 13g is closed. In this flow control mode, Q
Hydraulic pumps 1 and 2 so that 1 = Q1p and Q2 = Q2p
The discharge flow rate is controlled.

That is, for example, when the flow control command pressure Pq1 related to the flow control mechanism 11 increases, the force for pushing the flow control spool 11a to the left in the drawing increases, the flow control spring 11c is compressed, and the flow control spool 11a Move in the direction. Therefore, the control port 11e of the flow control feedback sleeve 11d is opened. When the control port 11e is opened, the large-diameter pressure chamber 3c of the differential cylinder 3
Are the actuator ports 7a and 11f and the control port 11
e, 13f, communicate with the tank 6 via the tank ports 13e, 7e, the oil in the large diameter side pressure chamber 3c is returned to the tank 6, the servo piston 3a is moved leftward in the figure, and the discharge flow rate of the hydraulic pump 1 is reduced. Increase.

When the servo piston 3a is moved in this manner, the movement of the servo piston 3a is transmitted to the feedback sleeve 11 by the feedback link 17.
d, the feedback sleeve 11d is moved following the flow control spool 11a, and when the feedback sleeve 11d moves the same amount as the flow control spool 11a, they become neutral, the servo piston 3a stops, and the hydraulic pump 1 The discharge flow rate is maintained.

When the flow control command pressure Pq1 becomes low, the force for pushing the flow control spool 11a to the left in the figure becomes weak, the flow control spring 11c is extended, and the flow control spool 11a is moved to the right in the figure. Therefore, the hydraulic pressure source port 11 of the flow control feedback sleeve 11d
g can be opened. When the hydraulic pressure source port 11g is opened, the large-diameter pressure chamber 3c of the differential cylinder 3
a, 11f and the hydraulic pressure source ports 11g, 7c, which communicate with the hydraulic pressure source 5, and the servo piston 3 is driven by the pressure receiving area difference between the large diameter side pressure chamber 3c and the small diameter side pressure chamber 3b of the differential cylinder 3.
a is moved rightward in the figure to decrease the discharge flow rate of the hydraulic pump 1.

When the servo piston 3a is moved in this manner, the movement of the servo piston 3a is transmitted to the feedback sleeve 11 by the feedback link 17.
d, the feedback sleeve 11d is moved to follow the flow control spool 11a, and when the feedback sleeve 11d moves by the same amount as the flow control spool 11a, the flow control spool 11a and the feedback sleeve 11d become neutral, and the servo piston 3a Stops, and the discharge flow rate of the hydraulic pump 1 is maintained.

Here, the neutral state of the flow control spool 11a and the feedback sleeve 11d means that the control port 1
1e and the hydraulic pressure source port 11g are respectively land portions 11a.
At 1 and 11a2, the aperture area is reduced to a minute aperture area, and the aperture state of both apertures is balanced. Flow control spool 12
The same applies to the neutral state of the feedback sleeve 12a and the horsepower control spool 13a and the feedback sleeve 13d.

FIG. 3 is a diagram showing the flow control characteristics in the flow control mode. The horizontal axis in FIG. 3 is the flow control command pressure Pq.
1 or Pq2, the vertical axis is the discharge flow rate Q1 of the hydraulic pumps 1 and 2
Or Q2. In the flow control mode, as described above, the discharge flow rate of the hydraulic pump 1 or 2 increases as the flow control command pressure Pq1 or Pq2 increases, and the discharge of the hydraulic pump 1 or 2 decreases as the flow control command pressure Pq1 or Pq2 decreases. The flow rate decreases. In the figure, Qmin is the minimum discharge flow rate.

<Horse Power Control Mode>
This is the case where Q1p + Q2p> 2Qc as described above. In this case, the horsepower control spool 13a and the feedback sleeve 13d of the horsepower control mechanism 13 shown in FIG. 1 are in a neutral state, and the flow control spools 11a and 12a and the flow control feedback sleeves 11d and 12d of the flow control mechanisms 11 and 12 are connected. The relative positional relationship between the control ports 11e,
12e is open, and the hydraulic pressure source ports 11g and 12g are closed. In this horsepower control mode, when Q1p ≧ Qc and Q2p ≧ Qc as described above, for example, Q1 = Qc
c, Q2 = Qc, and when Q1p> Qc, Q2p <Qc, Q1 = 2Qc−Q2 ≦ Qmax, Q2 = Q2p
When Q1p <Qc, Q2p> Qc, Q1 =
The discharge flow rates of the hydraulic pumps 1 and 2 are controlled such that Q1p, Q2 = 2Qc-Q1 ≦ Qmax.

First, the basic operation in the horsepower control mode will be described.

In the horsepower control mode, the flow control mechanism 11,
Even if the control ports 11e and 12e of the 12 are open, since the horsepower control spool 13a of the horsepower control mechanism 13 and the feedback sleeve 13d are in a neutral state, the servo pistons 3a and 4a of the differential cylinders 3 and 4 remain stopped. Therefore, the discharge flow rates of the hydraulic pumps 1 and 2 do not increase.

Further, for example, when the discharge pressure of the hydraulic pump 1 increases and the horsepower control command pressure Pd1 increases, the force for pushing the horsepower control spool 13a rightward in the drawing increases, and the horsepower control spring 13c1 (or 13c2) is compressed. Then, the horsepower control spool 13a is moved rightward. For this reason,
The hydraulic power source port 13g of the horsepower control feedback sleeve 13d is opened, and the large diameter side pressure chambers 3c, 4c of the differential cylinders 3, 4 are connected to the actuator ports 7a, 11f, the control ports 11e, 13f, the hydraulic power ports 13g, 7c or the actuator. Ports 7b and 12f, control port 12
e, 13f and the hydraulic pressure source 5 via the hydraulic pressure source ports 13g, 7d.
The servo pistons 3a, 4c are moved rightward in the figure by the pressure receiving area difference between the pressure chambers 4c and the small-diameter pressure chambers 3b, 4b to reduce the discharge flow rate of the hydraulic pumps 1, 2.

Then, as described above, the servo pistons 3a,
When the servo piston 3a is moved, the movement of the servo pistons 3a and 4a
9, the feedback sleeve 13d is moved following the horsepower control spool 13a, and when the feedback sleeve 13d moves by the same amount as the horsepower control spool 13a, they become neutral, and the servo piston 3a and 4a are stopped, and the discharge flow rate of the hydraulic pump 1 is maintained.

Further, the discharge pressure of the hydraulic pump 1 decreases,
When the horsepower control command pressure Pd1 decreases, the force for pushing the horsepower control spool 13a to the right in the figure decreases, the horsepower control spring 13c1 (or 13c2) is extended, and the horsepower control spool 13a moves to the left in the figure. Therefore, the tank port 13 of the horsepower control feedback sleeve 13d is
e is opened and the large-diameter pressure chamber 3 of the differential cylinders 3 and 4 is opened.
c and 4c are actuator ports 7a and 11f, control ports 11e and 13f, tank ports 13e or actuator ports 7b and 12f, and control ports 12e and 13
f, communicating with the tank 6 via the tank port 13e, the oil in the large-diameter pressure chambers 3c, 4c is returned to the tank 6, and the servo pistons 3a, 4a are moved leftward in the figure to discharge the hydraulic pumps 1, 2. Increase flow rate.

Then, as described above, the servo pistons 3a,
When the servo piston 3a is moved, the movement of the servo pistons 3a and 4a
9, the feedback sleeve 13d is moved to follow the horsepower control spool 13a. When the feedback sleeve 13d moves by the same amount as the horsepower control spool 13a, the feedback sleeve 13d becomes neutral and the servo piston 3a, 4a is stopped, and the discharge flow rates of the hydraulic pumps 1 and 2 are maintained.

The same applies when the discharge pressure of the hydraulic pump 2 changes in the horsepower control mode and when both of the discharge pressures of the hydraulic pumps 1 and 2 change.

Next, the horsepower control characteristics when Q1p ≧ Qc and Q2p ≧ Qc will be described.

FIG. 4 is a graph showing horsepower control characteristics when Q1p ≧ Qc and Q2p ≧ Qc. The horizontal axis in FIG. 4 is the average value (average discharge pressure) Pdm of the discharge pressures Pd1 and Pd2 of the hydraulic pumps 1 and 2 (Pdm = Pd1 + Pd2 / 2), and the vertical axis is the discharge flow rate Q1 or Q2 of the hydraulic pump 1 or 2. . H
Reference numerals 1 and H2 denote PQ lines for horsepower control, which are characteristics given by the horsepower control springs 13c1 and 13c2, respectively. Since Q1p ≧ Qc and Q2p ≧ Qc, Q1p,
Even if Q2p changes, the differential cylinders 3 and 4 do not operate, and H
The positions of 1, H2 do not change.

For example, in FIG. 4, it is assumed that the required flow rate of the hydraulic pump 1 (the required flow rate of the flow control command pressure Pq1) is Qt, and the average discharge pressure Pdm of the hydraulic pumps 1 and 2 is Pa. In this case, the hydraulic pump 1 operates at the point Ra on the PQ line H1, and the discharge flow rate of the hydraulic pump 1 is Qa (<
Qt). Thereafter, the average discharge pressure Pdm becomes P
b, the operating point of the hydraulic pump 1 becomes the PQ line H
1 to Rb, and the discharge flow rate of the hydraulic pump 1 becomes Q
b. As a result, the output horsepower of an engine (not shown) that rotationally drives the hydraulic pumps 1 and 2 becomes the PQ line H
1, the overload operation can be prevented without exceeding the limit value given by H2. Thereafter, when the average discharge pressure Pm decreases, the operating point of the hydraulic pump 1 moves on the PQ line H1 to the Ra side, and the discharge flow rate of the hydraulic pump 1 increases. When the average discharge pressure Pm further decreases to Pt, the discharge flow rate of the hydraulic pump 1 becomes Qt at the Rt point as the required flow rate. Thereafter, when the average discharge pressure Pm further decreases, the flow shifts to the flow control mode, and the discharge flow rate of the hydraulic pump 1 is maintained at Qt. The same applies to the hydraulic pump 2 side.

Next, variable horsepower control which is a feature of the present invention will be described.

In this embodiment, as described above, the two flow rate control mechanisms 11 and 12 and the horsepower control mechanism 13 are arranged at equal intervals, and the link mechanism including the links 17, 18 and 19 is the same as the link mechanism of the horsepower control mechanism 12. Feedback sleeve 13
The movement amount of d is configured to be the average value of the movement amounts of the feedback sleeves 11d and 12d of the flow control mechanisms 11 and 12. As a result, when considering the horsepower control in each of the hydraulic pumps 1 and 2, the position of the feedback sleeve 13d of the horsepower control mechanism 13 is not limited to the position of the feedback sleeve of the flow control mechanism related to itself, but also to the flow control mechanism of the other party. And the horsepower control characteristic is variable. That is, as described above, Q1p> Qc, Q2p
<Qc, Q1 = 2Qc−Q2 ≦ Qmax, Q2
= Q2p, and when Q1p <Qc, Q2p> Qc, Q1 = Q1p and Q2 = 2Qc−Q1 ≦ Qmax.

FIG. 5 is a diagram for explaining the variable horsepower control characteristics. FIG. 5 (a) shows the hydraulic pump 1 side, and FIG. 5 (b) shows the hydraulic pump 2 side. The horizontal axis and the vertical axis in FIGS. 5A and 5B are the average discharge pressure Pdm and the discharge flow rates Q1 and Q2 of the hydraulic pumps 1 and 2 as in FIG. FIG.
In (a) and (b), X1 and Y1 are characteristics of the average discharge flow rate Qc during the horsepower control of the hydraulic pumps 1 and 2.
That is, during the horsepower control, the average discharge flow rate Qc of the hydraulic pumps 1 and 2 changes with respect to the average discharge pressure Pdm as X1, Y1.

First, when Q1p ≧ Qc and Q2p ≧ Qc, assuming that the discharge flow rates Q1 and Q2 of the hydraulic pumps 1 and 2 are controlled to be equal, the horsepower control characteristic is the same as the horsepower control characteristic of the average discharge flow Qc. X1 and Y1 are the same. At this time, for example, the average discharge pressure Pdm of the hydraulic pumps 1 and 2
Is Pc, the discharge flow rates Q1 and Q2 of the hydraulic pumps 1 and 2 at that time are Qco on the horsepower control characteristics X1 and Y1.
(Points E1 and F1). Further, as described above, when the average discharge pressure Pdm changes, the operating points of the hydraulic pumps 1 and 2 change on X1 and Y1, and the discharge flow rate increases and decreases accordingly.

Next, from that state, if the hydraulic pump 2
Required flow rate Q2p is ΔQa lower than Qc (Q <
c), the hydraulic supply port 12g of the flow control sleeve 12d is opened, so that the actual discharge flow rate Q2 of the hydraulic pump 2 is reduced.
Is reduced to Q2pa (point F2), and the feedback sleeve 13d of the horsepower control mechanism 13 is moved rightward in FIG. As a result, the tank port 13e of the feedback sleeve 13d of the horsepower control mechanism 13 is opened to allow a margin for horsepower control, and the discharge flow rate Q1 of the hydraulic pump 1 can be increased by ΔQa. For this reason, if the required flow rate Q1p of the hydraulic pump 1 is equal to or more than Qc + ΔQa, the actual discharge flow rate Q1 of the hydraulic pump 1 increases to Q1pa of Qc + ΔQa (point E2).

When the required flow rate Q2p of the hydraulic pump 2 further decreases and drops to Q2pb smaller than Qc by ΔQb (= Qmax−Qc), the hydraulic source port 12g of the flow rate control sleeve 12d is opened again, and the actual hydraulic pump 2 is operated. Is reduced to Q2pb (point F3), and the feedback sleeve 13d of the horsepower control mechanism 13 is further moved rightward in the figure. As a result, the horsepower control of the horsepower control mechanism 13 has a margin again, and the discharge flow rate Q1 of the hydraulic pump 1 becomes ΔQb
Can only be increased. Therefore, if the required flow rate Q1p of the hydraulic pump 1 is equal to or more than Qc + ΔQb, the actual discharge flow rate Q1 of the hydraulic pump 1 is Qma of Qc + ΔQb.
x (point E3).

Here, Qmax in FIG. 5 is the mechanical upper limit of the hydraulic pump 2, and thereafter, the required flow rate Q
Even if 2p is further reduced and the actual discharge flow rate Q2 is further reduced, the discharge flow rate Q1 of the hydraulic pump 1 does not increase beyond Qmax.

The above description is for the case where the required flow rate of the hydraulic pump 2 is reduced. The same applies to the case where the required flow rate of the hydraulic pump 1 is reduced. The horsepower control characteristic of the hydraulic pump 1 is lower than X1. Side (flow decreasing side), and the horsepower control characteristic of the hydraulic pump 2 side changes upward from Y1 (flow increasing side).

As described above, in the horsepower control of the present embodiment, when Q1p + Q2p> 2Qc, Q1p> Qc, Q2p <
In the case of Qc, Q1 = 2Qc−Q2 ≦ Qmax, Q2 =
Q2p, and when Q1p <Qc, Q2p> Qc,
Q1 = Q1p, Q2 = 2Qc-Q1 ≦ Qmax.

In other words, the required flow rate Q2 of the hydraulic pump 2
When p is Q2p <Qc, the discharge flow rate of the hydraulic pump 1 can be increased to Qc + (Qc−Q2p) = 2Qc−Q2p, and conversely, the required flow rate Q1p of the hydraulic pump 1 becomes Q1p
When <Qc, the discharge flow rate of the hydraulic pump 2 is Qc +
(Qc-Q1p) = 2Qc-Q1p. That is, when the required flow rate of the other hydraulic pumps is smaller than the average discharge flow rate Qc of both pumps at the time of horsepower control, the discharge flow rates of the hydraulic pumps 1 and 2 can be greater than the average flow discharge rate Qc. As a result, the output power of the engine as the power source can be effectively used.

Next, an actual operation example will be described.

<Hydraulic Pumps 1 and 2 Standby> FIG.
FIG. 5 is a diagram illustrating an operation state of the hydraulic pump displacement control device when both of the hydraulic pumps 1 and 2 are in a standby state. In this standby state, the flow control command pressure Pq
1 and Pq2 are at the minimum pressure Pqmin, and the spools 11a and 12a of the flow control mechanisms 11 and 12 are at the initial positions on the right side in the figure. Also, since the tank port 13e of the feedback sleeve 13d of the horsepower control mechanism 13 is open, the servo pistons 3a, 4a of the differential cylinders 3, 4 and the flow control feedback sleeves 11d, 12d also follow the spools 11a, 12a. In the initial position on the right side, the discharge flow rates of the hydraulic pumps 1 and 2 are kept at the minimum standby flow rate Qmin. The state of both pumps at this time is indicated by point A in the flow control characteristic diagram of FIG.

<Hydraulic Pump 1: Flow Control Mode, Hydraulic Pump 2: Standby> FIG. 7 shows the operating state of the hydraulic pump displacement control device when the hydraulic pump 1 is in the flow control mode and the hydraulic pump 2 is in the standby state. FIG.
In this state, the spool 11a of the flow control mechanism 11 shifts leftward in the figure due to the flow rate command pressure Pq1, and accordingly, the servo piston 3a of the differential cylinder 3 and the feedback sleeve 11d also shift leftward in the figure. 1 is controlled to be a discharge flow rate corresponding to the required flow rate of the flow rate command pressure Pq1. The hydraulic pump 1 at this time
Is shown at point B in the flow control characteristic diagram of FIG.

<Flow Control Mode for Hydraulic Pumps 1 and 2> FIG. 8 is a diagram showing an operation state of the hydraulic pump displacement control device when both the hydraulic pumps 1 and 2 are in the flow control mode. In this state, the spool 12 of the flow control mechanism 12
a is also shifted leftward in the figure by the flow rate command pressure Pq2,
In response to this, the servo piston 4a of the differential cylinder 4 and the feedback sleeve 12d also shift to the left in the drawing, and the hydraulic pump 2 is also controlled to discharge a flow rate corresponding to the required flow rate of the flow rate command pressure Pq2. The hydraulic pump 2 at this time is also located, for example, at point B in the flow control characteristic diagram of FIG.

<Hydraulic Pump 1: Divergence Point of Horsepower Control and Flow Control, Hydraulic Pump 2: Standby> FIG. 9 is at a branch point where the hydraulic pump 1 shifts from the flow control mode to the horsepower control mode. FIG. 4 is a diagram illustrating an operation state of the hydraulic pump displacement control device in a standby state. This state is a case where the horsepower control command pressure Pd1 has risen from the state of FIG. 7, the spool 13a of the horsepower control mechanism 13 shifts rightward in the drawing, and the tank port 13e of the horsepower control feedback spool 13d has a small opening. It has reached a neutral state with a limited area.

FIG. 10 shows the horsepower control characteristics at this time.
In FIG. 10, Pac is an average discharge pressure of the hydraulic pumps 1 and 2. Since the hydraulic pump 2 is in the standby state, the discharge pressure of the hydraulic pump 1 is approximately 2 Pac. Qac
Is the average discharge flow rate of the hydraulic pumps 1 and 2 at the time of horsepower control corresponding to Pac. At this time, since the hydraulic pump 2 is in the standby state, the discharge flow rate of the hydraulic pump 1 is approximately 2Qa.
c. Q1p is a required flow rate of the hydraulic pump 1, and this required flow rate Q1p is substantially equal to 2Qac, and the operating point of the hydraulic pump 1 is at a point C. That is, the discharge flow rate of the hydraulic pump 1 is controlled to be 2Qac. Point D indicates the standby state of the hydraulic pump 2 at this time.

<Hydraulic pump 1: Start → end of horsepower control → Hydraulic pump 2: Standby> FIG. 11 shows a hydraulic pump when the discharge pressure of the hydraulic pump 1 increases from the state of FIG. 9 and the horsepower control of the hydraulic pump 1 starts. It is a figure showing the operation state of a capacity control device. In this state, the spool 13a of the horsepower control mechanism 13 further shifts rightward in the figure due to the increase in the horsepower control command pressure Pd1, and the horsepower control feedback sleeve 13
Since the hydraulic pressure source port 13g of FIG.
Discharge flow rate decreases.

FIG. 12 shows a state after the discharge flow rate of the hydraulic pump 1 has decreased. In this state, the horsepower control sleeve 13d moves following the spool 13a, and the hydraulic power source port 13g returns to the neutral state in which the opening area is reduced to a small opening area again.

FIG. 13 shows the horsepower control characteristics at this time.
In FIG. 13, Pbc is the average discharge pressure after the discharge pressure of the hydraulic pump 1 has increased, and Qbc is the average discharge flow rate during horsepower control corresponding to Pbc. The upper limit of the discharge flow rate of the hydraulic pump 1 is reduced from 2Qac to 2Qbc. Therefore, the operating point of the hydraulic pump 1 moves from the point C to the point E, and the discharge flow rate is reduced from 2Qac to 2Qbc. At this time, the hydraulic pump 2 is in a standby state at point F.

<Hydraulic Pumps 1, 2: Branching Point of Flow Control and Horsepower Control> FIG. 14 shows the hydraulic pressure when both of the hydraulic pumps 1, 2 are at the branching point at which the mode is simultaneously shifted from the flow control mode to the horsepower control mode. It is a figure showing an operating state of a pump displacement control device. This state is when the horsepower control command pressure Pd1 and / or Pd2 has risen from the state shown in FIG. 8, and the spool 13a of the horsepower control mechanism 13 shifts rightward in the drawing, and the tank port of the horsepower control feedback spool 13d 13e has reached a neutral state in which it is narrowed to a small opening area.

FIG. 15 shows the horsepower control characteristics at this time.
In FIG. 15, Pcc is the average discharge pressure of the hydraulic pumps 1 and 2, and Qcc is the average discharge flow rate of the hydraulic pumps 1 and 2 corresponding to Pcc during horsepower control. Being at a branch point between the flow control and the horsepower control means that the required flow rates Q1p and Q2p of the hydraulic pumps 1 and 2 are both substantially equal to Qcc. Therefore, the horsepower control characteristics of the hydraulic pumps 1 and 2 are X1 and Y1, which are the same as the characteristics of the average discharge flow rate, and the operating points of the hydraulic pumps 1 and 2 are at points G and H on X1 and Y1. That is, the discharge flow rates of the hydraulic pumps 1 and 2 are each controlled to be Qcc.

<Hydraulic pumps 1 and 2: Power control start → end> FIG. 16 shows that the average discharge pressure increases from the state of FIG.
FIG. 5 is a diagram illustrating an operation state of the hydraulic pump displacement control device when horsepower control of the hydraulic pumps 1 and 2 starts. In this state, when the horsepower control command pressure Pd1 and / or Pd2 further increases, the spool 13a of the horsepower control mechanism 13 further shifts rightward in the drawing, and the horsepower control feedback sleeve 13
Since the hydraulic pressure source port 13g of d is opened, the discharge flow rate of the hydraulic pumps 1 and 2 further decreases.

FIG. 17 shows a state after the discharge flow rates of the hydraulic pumps 1 and 2 have been reduced. In this state, the horsepower control sleeve 13d moves following the spool 13a, and the hydraulic power source port 13g returns to the neutral state in which the opening area is reduced to a small opening area again.

FIG. 18 shows the horsepower control characteristics at this time.
In FIG. 18, Pdc is the average discharge pressure after the rise, and Qdc is the average discharge flow rate during horsepower control corresponding to Pdc. Required flow rates Q1p, Q2p of hydraulic pumps 1, 2
Does not change with Qcc (> Qdc), the horsepower control characteristics of the hydraulic pumps 1 and 2 do not change with X1 and Y1. Therefore, the operating points of the hydraulic pumps 1 and 2 move from point G to point I and from point H to point J on X1 and Y1, respectively, and the discharge flow rate decreases from Qcc to Qdc, respectively.

<Hydraulic pumps 1 and 2: Branch point between flow control and horsepower control> FIG. 19 shows a branch point where the hydraulic pumps 1 and 2 simultaneously shift from the flow control mode to the horsepower control mode similarly to FIG. FIG. 5 is a diagram showing an operation state of the hydraulic pump displacement control device of FIG. However, the average discharge pressure of the hydraulic pumps 1 and 2 is higher than in the case of FIG. 14, and the spool 13a of the horsepower control mechanism 13 is further shifted rightward in the figure. FIG. 20 is a horsepower control characteristic diagram at this time, wherein the average discharge pressure of the hydraulic pumps 1 and 2 is Pec (> Pcc), and the required flow rates Q1p and Q2p of the hydraulic pumps 1 and 2 are respectively the average of the horsepower control. The operating points of the hydraulic pumps 1 and 2 are substantially equal to the discharge flow rate Qec, and the operating points of the characteristics X1 and Y1
At the upper points K and L, the discharge flow rates of the hydraulic pumps 1 and 2 are controlled to be Qec.

<Hydraulic pumps 1, 2: horsepower control> FIG.
20 is a diagram showing an operation state of the hydraulic pump displacement control device when the required flow rate of the hydraulic pump 1 has increased from the state shown in FIG. In this state, as the flow control command pressure Pq1 increases, the spool 11a of the flow control mechanism 11 moves leftward in the drawing, and the control port 11e of the flow control feedback sleeve 11d is opened. However, since the horsepower control mechanism 13 is in the horsepower control mode, the position of the servo piston 3a of the differential cylinder 3 is maintained, and the discharge flow rate of the hydraulic pump 1 does not increase.

FIG. 22 shows the horsepower control characteristics at this time.
In FIG. 22, the required flow rate Q1p of the hydraulic pump 1 is shown in FIG.
0 compared to Q1p. However, since the horsepower control characteristics of the hydraulic pumps 1 and 2 remain unchanged at X1 and Y1, the operating points of the hydraulic pumps 1 and 2 are also the K points on X1 and Y1,
At the point L, the discharge flow rate of the hydraulic pump 1 is Qec
Is held in.

<Hydraulic pumps 1, 2: Variable horsepower control> FIG.
FIG. 3 is a diagram illustrating an operation state of the hydraulic pump displacement control device when the required flow rate of the hydraulic pump 2 decreases from the state illustrated in FIG. 21. That is, when the flow control command pressure Pq1 decreases in the state of FIG. 21, the spool 12a of the flow control mechanism 12
Moves rightward in the figure and opens the hydraulic source port 12g of the flow control feedback sleeve 12d, so that the servo piston 4a of the differential cylinder 4 is moved rightward in the figure, and the discharge flow rate of the hydraulic pump 2 decreases. The movement of the servo piston 4a is transmitted to the flow control sleeve 12d, and the flow control sleeve 12d moves rightward in the figure following the movement, and the hydraulic power source port 12g of the feedback sleeve 12d is narrowed to a small opening area. When the neutral state is reached, the servo piston 4a stops, and the discharge flow rate of the hydraulic pump 2 at that time is maintained. This movement of the feedback sleeve 12d is also transmitted to the feedback sleeve 13d of the horsepower control mechanism 13, and the tank port 13e of the horsepower control sleeve 13d is opened to similarly move the feedback sleeve 13d rightward in the drawing. The large diameter side pressure chamber 3c communicates with the tank. Therefore, the servo piston 3a moves to the left in the figure, and the discharge flow rate of the hydraulic pump 1 increases. The movement of the servo piston 3a is transmitted to the horsepower control sleeve 13d via the flow control sleeve 11d, and the horsepower control sleeve 13d moves leftward in the drawing following the movement, and the tank port 13e of the feedback sleeve 13d has a small opening area. When the servo piston 3a stops when the neutral state is reached, the discharge flow rate of the hydraulic pump 1 at that time is maintained.

FIG. 24 shows the horsepower control characteristics at this time.
In FIG. 24, the required flow rate Q2p of the hydraulic pump 2 is shown in FIG.
2, the horsepower control characteristic of the hydraulic pump 2 changes, and the operating point of the hydraulic pump 2 moves from the L point to the N point. Correspondingly, the horsepower control characteristic of the hydraulic pump 1 changes, the operating point of the hydraulic pump 1 also moves from the point K to the point M, and the discharge flow rate of the hydraulic pump 1 changes from Qec to Q1p.
e.

As described above, according to the present embodiment, a common horsepower control mechanism is provided for the two hydraulic pumps 1 and 2, so that the structure of the regulator 7 of the hydraulic pump displacement control device is simplified, and The size can be reduced.

The position of the feedback sleeve of the horsepower control mechanism is changed not only by the change of the pump discharge flow rate related to itself but also by the change of the pump discharge flow rate of the partner side by the link mechanism including the links 17, 18, and 19. , The horsepower control characteristic can be changed according to the pump discharge flow rate of the other party, whereby the horsepower control limit value Qc
If there is room for the pump discharge flow on the other side,
The pump discharge flow rate of the pump itself can be increased by the flow amount corresponding to the margin, so that the output horsepower of the engine can be effectively used to the maximum.

In the above embodiment, the case where there are two hydraulic pumps has been described. However, the present invention can be similarly applied to a case where there are three or more hydraulic pumps. For example, if there are three hydraulic pumps, three flow control mechanisms are arranged so that the spool center axis is located at the top of the triangle, and the horsepower control mechanism is located between the three flow control mechanisms (center of the triangle). What is necessary is just to arrange so that it may be located.
Also in this case, by linking the feedback sleeve of each flow control mechanism, the feedback sleeve of the horsepower control mechanism, and the differential cylinder of the three hydraulic pumps with a link mechanism, the structure can be simplified and downsized, and the limit value of the horsepower control can be achieved. In contrast, when there is a margin in the other party's pump discharge flow rate, the own pump discharge flow rate can be increased by the surplus flow rate, and the output horsepower of the engine can be used most effectively.

[0089]

According to the present invention, the regulator structure of the hydraulic pump displacement control device can be simplified and the size can be reduced.

Further, when there is a margin in the pump discharge flow rate on the partner side with respect to the limit value of the horsepower control, the pump discharge flow rate of the own pump can be increased by the marginal flow rate, and the output horsepower of the engine is maximized. It can be used effectively.

[Brief description of the drawings]

FIG. 1 is a diagram showing a structure of a hydraulic pump displacement control device according to one embodiment of the present invention.

FIG. 2 is a diagram showing the hydraulic pump displacement control device shown in FIG. 1 in a hydraulic circuit.

FIG. 3 is a view showing a flow control characteristic in a flow control mode of the hydraulic pump displacement control device shown in FIG. 1;

FIG. 4 is a diagram showing horsepower control characteristics of the hydraulic pump displacement control device shown in FIG.

FIG. 5 is a view showing a variable horsepower control characteristic of the hydraulic pump displacement control device shown in FIG. 1;

FIG. 6 is a diagram illustrating an operation state of the hydraulic pump displacement control device when both of the two hydraulic pumps are in a standby state.

FIG. 7 is a diagram showing an operation state of the hydraulic pump displacement control device when one hydraulic pump is in a flow control mode and the other hydraulic pump is in a standby state.

FIG. 8 is a diagram showing an operation state of the hydraulic pump displacement control device when both hydraulic pumps are in the flow control mode.

FIG. 9 is a diagram showing an operation state of the hydraulic pump displacement control device when one hydraulic pump is at a branch point where the flow rate control mode shifts to the horsepower control mode and the other hydraulic pump is in a standby state.

FIG. 10 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG. 9;

11 is a diagram showing an operation state of the hydraulic pump displacement control device when the discharge pressure of one of the hydraulic pumps increases from the state of FIG. 9 to start horsepower control.

12 is a diagram showing an operation state of the hydraulic pump displacement control device after the discharge flow rate of one hydraulic pump has decreased in the state of FIG.

13 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG.

FIG. 14 is a diagram showing an operation state of the hydraulic pump displacement control device when both of the two hydraulic pumps are at a branch point where a transition from the flow control mode to the horsepower control mode is made at the same time.

FIG. 15 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG.

FIG. 16 shows that the average discharge pressure increases from the state of FIG.
FIG. 5 is a diagram showing an operation state of a hydraulic pump displacement control device when horsepower control of two hydraulic pumps starts.

FIG. 17 is a diagram showing an operation state of the hydraulic pump displacement control device after the discharge flow rates of the two hydraulic pumps have decreased in the state shown in FIG. 16;

18 is a diagram showing the horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG.

FIG. 19 is a diagram illustrating an operation state of the hydraulic pump displacement control device at a branch point where two hydraulic pumps simultaneously shift from the flow control mode to the horsepower control mode.

20 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG. 19;

FIG. 21 is a diagram showing an operation state of the hydraulic pump displacement control device when the required flow rate of one hydraulic pump increases from the state shown in FIG. 19;

FIG. 22 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG. 21;

FIG. 23 is a diagram showing an operation state of the hydraulic pump displacement control device when the required flow rate of the other hydraulic pump has decreased from the state shown in FIG. 21;

24 is a diagram showing horsepower control characteristics when the hydraulic pump displacement control device is in the state shown in FIG. 23.

FIG. 25 is a view showing the structure of a conventional hydraulic pump displacement control device.

FIG. 26 is a diagram showing a conventional hydraulic pump displacement control device in a hydraulic circuit.

FIG. 27 is a diagram showing horsepower control characteristics of a conventional hydraulic pump displacement control device.

[Explanation of symbols]

 1, 2 Hydraulic pump 1a, 2a Variable displacement mechanism 3, 4 Differential cylinder 3a, 4a Servo piston 3b, 4b Small-diameter pressure chamber 3c, 4c Large-diameter pressure chamber 5 Hydraulic source 6 Tank 7 Regulator 7a, 7b Actuator port 7c, 7d Hydraulic source port 7e Tank port 7f, 7g Flow control port 7h, 7i Horsepower control port 11, 12 Flow control mechanism 11a, 12a Flow control spool 11b, 12b Command pressure receiving part 11c, 12c Flow control spring 11d, 12d Feedback Sleeve 11e, 12e Control port 11f, 12f Actuator port 11g, 12g Hydraulic power source port 13 Horsepower control mechanism 13a Horsepower control spool 13b1, 13b2 Command pressure receiving part 13c1, 13c2 Horsepower control spring 13d Feedback three 13e tank port 13f control port 13g hydraulic power source port 15 casing 17, 18 feedback link (link mechanism) 19 intermediate link (link mechanism) 20, 21 internal passages 22, 23 internal passage

Claims (3)

[Claims] At least two differential cylinders for respectively operating variable displacement mechanisms of at least two variable displacement hydraulic pumps driven by one power source, and a pressure between the differential cylinder, the hydraulic source and the tank. A regulator for controlling oil supply / discharge; the regulator includes a flow control unit that operates according to a flow control command pressure, and a horsepower control unit that operates according to a horsepower control command pressure. Means for controlling the servo pistons of said two differential cylinders by means and controlling the displacement variable mechanism of said two hydraulic pumps, wherein said flow rate control means comprises: Having at least two flow control mechanisms for inputting a control command pressure and individually operating the two differential cylinders, A hydraulic pump displacement control device, wherein the horsepower control means has a common horsepower control mechanism for inputting respective horsepower control command pressures of the at least two hydraulic pumps and operating the two differential cylinders. 2. The hydraulic pump displacement control device according to claim 1, wherein said at least two flow control mechanisms and said common horsepower control mechanism link respective feedback sleeves to each other and connect said feedback sleeves to each other. A hydraulic pump displacement control device, further comprising a link mechanism for linking with a servo piston of a cylinder. 3. The hydraulic pump displacement control device according to claim 2, wherein said common horsepower control mechanism is disposed between said at least two flow control mechanisms, and said link mechanism is a feedback of said common horsepower control mechanism. A hydraulic pump displacement control device characterized in that the movement amount of the sleeve is configured to be an average value of the movement amounts of the feedback sleeves of the at least two flow control mechanisms.