WO2014091690A1

WO2014091690A1 - Variable displacement pump regulator

Info

Publication number: WO2014091690A1
Application number: PCT/JP2013/006898
Authority: WO
Inventors: 隆二堺; 勇吉村; 亮次入江; 秀俊三浦
Original assignee: 川崎重工業株式会社
Priority date: 2012-12-11
Filing date: 2013-11-25
Publication date: 2014-06-19
Also published as: EP2933490A4; EP2933490A1; JP2014114911A; CN104411973B; KR20150003366A; US20150226190A1; JP5918688B2; CN104411973A

Abstract

This regulator (10) is provided with a feedback lever (24) which moves a compensator sleeve (32) in the axial direction of a compensator spool (30). The compensator spool (30) and the compensator sleeve (32) form an autogenous pressure introducing chamber (34), into which the autogenous pressure (Pd) of a pump is introduced, and a pressure adjusting unit (51), which outputs a control pressure (Pcl) from the autogenous pressure introducing chamber (34) to a large-diameter pressure receiving part (22) of a servo piston (21). Inside of the compensator sleeve (32), a step portion (42) is provided where a pressure acts in the direction a control spring (31) biases the compensator spool (30) by means of the autogenous pressure (Pd) introduced into the autogenous pressure introducing chamber (34).

Description

Variable displacement pump regulator

The present invention relates to a variable displacement pump regulator capable of controlling a discharge flow rate by adjusting a tilt angle.

Conventionally, power machines such as hydraulic excavators, cranes, wheel loaders, bulldozers, etc. (in this specification and claims, these power machines (heavy machinery) are collectively referred to as “work machines”) Used for construction work. For example, a hydraulic excavator will be described as an example. In a hydraulic excavator, various actuators are used to rotate the upper swing body and operate buckets, arms, booms, and the like, and hydraulic pressure is used to drive these actuators. Yes.

Such actuators are supplied with hydraulic oil at a required flow rate from a hydraulic pump according to, for example, the turning speed in the case of an upper turning body and the weight of the arm or boom that is pumped up by a bucket. . In addition, a tandem hydraulic pump including a plurality of hydraulic pumps is used according to the number of actuators and necessary power.

¡A variable displacement pump is used for such a hydraulic pump, and the displacement of the variable displacement pump is controlled by a regulator. For example, in the case of using a pair of variable displacement pumps, each variable displacement pump is provided with a regulator, and each regulator does not exceed the horsepower of the engine that drives the pumps of both pumps. Thus, the total horsepower control for controlling the tilt angle of each pump is performed based on the discharge pressures of both pumps. In a general regulator, the position of the spool or sleeve is controlled by a signal pressure such as the self-pressure of the pump (the discharge pressure of the pump provided with the regulator), and the pressure oil having a pressure corresponding to the position is regulated. The pressure oil having a control pressure is output from the pressure adjusting unit and the capacity of the hydraulic pump is adjusted. In a hydraulic pump using a servo piston as one part of the capacity adjustment mechanism, the capacity of the variable displacement pump is adjusted by the position of the servo piston. The capacity adjustment in the swash plate type variable displacement pump is performed by the tilt angle of the swash plate.

For example, as this type of prior art, two variable displacement pumps driven by a prime mover, a plurality of actuators driven by pressure oil supplied from the variable displacement pump, and commands to a plurality of actuators The first detection means for detecting the command signal of the operation command means, the second detection means for detecting the loads of the plurality of actuators, and the input means for instructing the reference target rotational speed of the prime mover, the reference target rotational speed is low. There is an auto accelerator device that calculates a reference width for rotational speed correction that becomes smaller as it goes, and corrects a correction value of a reference target rotational speed by a correction value correcting means (see, for example, Patent Document 1).

In this auto accelerator device, the capacity of two variable displacement pumps is controlled by two regulators. Each regulator is guided by the self-pressure of the variable displacement pump provided with the regulator and the other pressure of the other variable displacement pump, and the position of the spool is controlled by the operation piston operating at those pressures. .

Japanese Patent Laid-Open No. 11-107322

By the way, in the regulator of the variable displacement pump as described above, the servo piston is moved by outputting the control pressure from the pressure adjusting unit according to the signal pressure such as the discharge pressure (self pressure) of the pump acting on the spool. The pump capacity is changed. When the servo piston moves, the regulator sleeve is also moved by the feedback lever engaged with the servo piston, and the tilt angle of the pump is controlled to be a target value. The spool position control by such self-pressure and the sleeve position control through the feedback lever are always performed, and the control pressure from the regulator of the regulator is guided to the servo piston by these, and the hydraulic pump Is controlled in accordance with the load of the actuator of the work machine.

However, an engaging pin is provided at the engaging portion between the feedback lever and the sleeve described above, and the contact portion of the engaging pin may be worn out over time.

In the case of the regulator of the auto accelerator device, since the operation piston for controlling the position of the spool is separate from the spool, the urging force in the axial direction does not act on the sleeve, and the engagement portion contacting the engagement pin does not A frictional force or the like is applied regardless of the operation direction of the spool, and the engaging portion is worn without directionality by the force. This wear causes a gap between the engagement pin and the contact portion, and this gap creates a dead band of a predetermined width with respect to the control pressure of the regulator that increases or decreases the flow rate of the pump. Causes non-directional hysteresis. This hysteresis increases with the progress of wear of the engaging portion due to use over time.

FIG. 8 is a diagram showing the horsepower control characteristic by the regulator of the hydraulic pump in relation to the discharge pressure and the flow rate, and the control line 101 is set so as to approximate the design horsepower line 100. This control line 101 shows an example in which two control springs are used, and the control line is changed by changing the number of control springs used for control during the horsepower control from one to two. Is changed so that the control line 101 approximates the equal horsepower line 100.

Therefore, as described above, when the engaging portion between the sleeve and the pin is worn without directionality, as shown in the figure, hysteresis occurs in the actual control line 102 with respect to the target flow rate of the control line 101, In some cases, the control line may swing to either the flow rate decrease side or the flow rate increase side.

When the flow rate with respect to the discharge pressure exceeds the flow rate of the equal horsepower line 100, for example, in the case of a hydraulic excavator, when performing work requiring large horsepower, such as excavation and loading work of earth and sand, As a result of discharging an excessive discharge flow rate, the horsepower of the hydraulic pump (or the total horsepower of all the hydraulic pumps if there are multiple hydraulic pumps) may exceed the horsepower of the prime mover, which may cause the prime mover to stall or become unstable There is.

Thus, when the hysteresis of the flow rate with respect to the discharge pressure of the pump increases due to wear at the engagement portion between the sleeve and the feedback lever, the control line 101 tries to lower the horsepower of the pump when the horsepower control of the regulator functions. As described above, even if the discharge flow rate of the pump is decreased, the discharge flow rate is increased by the hysteresis, and the possibility of causing a stall of the prime mover is increased.

Therefore, an object of the present invention is to provide a variable displacement pump regulator that can perform stable horsepower control even when flow rate hysteresis occurs in the horsepower control characteristics of a hydraulic pump over time.

To achieve the above object, a variable displacement pump regulator according to the present invention includes a servo piston that changes a discharge flow rate of a variable displacement pump driven by a prime mover, a feedback lever that detects the position of the servo piston, A compensator spool biased in one direction by a control spring, and the compensator spool is positioned around the compensator spool and is engaged with the feedback lever by the action of the servo piston and the feedback lever. A compensation sleeve configured to move in the axial direction, a discharge pressure introduction chamber into which the discharge pressure of the pump is introduced, and a control pressure from the discharge pressure introduction chamber to the large-diameter pressure receiving portion of the servo piston. The pressure adjusting unit to output is based on the compensation spool and the compensation sleeve. Formed Te, the compensator sleeve, the step portion acting pressure on the end direction by a discharge pressure introduced into the discharge pressure introducing chamber.

In this configuration, since the step portion is provided in the compensator sleeve that moves in the axial direction around the compensator spool, the compensator sleeve is attached in the opposite direction (one end direction) to the control spring by the discharge pressure introduced into the discharge pressure introduction chamber. Be forced. As a result, a force is always applied to the engaging portion of the feedback lever engaged with the compensator sleeve in the direction opposite to the control spring, and the wear portion of the engaging portion is limited. As a result, even if the control spring side of the engaging portion between the feedback lever and the compensator is worn due to aging, the horsepower control characteristic of the pump will decrease the discharge flow rate of the pump. Therefore, even if the engagement part of the feedback lever is worn, the horsepower of the pump does not exceed the horsepower of the prime mover (in some cases, the share of the single pump in the prime mover's horsepower), and stable horsepower control is performed to stall the prime mover. Can be prevented.

Further, the stepped portion may be formed on the opposite side of the discharge pressure introducing chamber from the pressure adjusting portion by reducing the inner diameter of the compensator sleeve.

With this configuration, the pressure adjusting unit side and the counter pressure adjusting unit of the compensator sleeve are biased in the opposite direction (one end direction) to the control spring by the discharge pressure introduced into the discharge pressure introducing chamber. A pressure receiving area difference can be provided on the side.

The discharge pressure introduction chamber is a self-pressure introduction chamber into which a self-pressure that is a discharge pressure of the pump is introduced, and is located at a position away from the self-pressure introduction chamber on the side opposite to the pressure regulating unit. A counterpart pressure introduction chamber into which a counterpart pressure that is a discharge pressure of the variable displacement pump is introduced may be formed by the compensation spool and the compensation sleep. “Self-pressure” in this specification and claims refers to the discharge pressure of its own variable displacement pump in a configuration having a plurality of pumps, and “other pressure” refers to another variable displacement pump. Discharge pressure.

If comprised in this way, in the structure provided with the several variable displacement pump, according to each discharge pressure, the total horsepower of the pump provided with the said regulator and other pumps may not exceed the horsepower of a prime mover. The pump capacity can be changed with a regulator.

The counterpart pressure introduction chamber is formed to have a smaller diameter than the self pressure introduction chamber, and the compensator has a step portion on which the pressure acts in the one end direction by the counterpart pressure introduced into the counterpart pressure introduction chamber. It may be provided.

With this configuration, even in a configuration including a plurality of variable displacement pumps, in which the pump displacement is controlled according to the self pressure and the counterpart pressure, the compensator is always in the direction opposite to the control spring (one end direction). It is possible to apply a force for urging. In other words, the force that always urges the engagement portion of the feedback lever toward the side where the pump capacity decreases can be applied by the counterpart pressure introduced into the counterpart pressure introduction chamber.

According to the present invention, since wear or the like occurs only on the side where the capacity of the pump is reduced even when used over time, it is possible to configure a variable capacity pump regulator capable of stable horsepower control even over time. Become.

FIG. 1 is a hydraulic circuit diagram showing an embodiment of a variable displacement pump regulator according to the present invention. 2 is a cross-sectional view of the variable displacement pump regulator shown in FIG. 1, and is a cross-sectional view taken along the line II-II shown in FIG. 3 is a cross-sectional view taken along arrows III-III shown in FIG. 4 is a cross-sectional view showing a state in which the compensator spool of the variable displacement pump regulator shown in FIG. 2 has moved to the control spring side. FIG. 5 is a cross-sectional view showing a state in which the compensator has changed in capacity from the state in which the compensator shown in FIG. 4 has moved toward the control spring, and the capacity of the pump has changed. FIG. 6 is a schematic diagram showing the relationship between the discharge pressure acting on the compensator shown in FIG. 2 and the force acting on the engaging portion of the feedback lever. FIG. 7 is a horsepower control diagram showing control characteristic changes that may occur when the variable displacement pump regulator shown in FIG. 2 changes over time. FIG. 8 is a horsepower control diagram showing a control characteristic change that may occur when the conventional variable displacement pump regulator changes over time.

Hereinafter, an embodiment of the present invention will be described with reference to the drawings. In the following embodiment, only one pump of a swash plate type double pump such as a tandem pump is illustrated, and the discharge pressure of the illustrated pump is “self pressure Pd”, and the discharge pressures of the other pumps are “other pressure P2”. " Further, only the portion related to the horsepower control of the variable displacement pump will be illustrated and described. Furthermore, the concept of the up / down / left / right direction in the document of this specification and claims shall be the same as the concept of the up / down / left / right direction in the sectional view of the regulator shown in FIG.

As shown in FIG. 1, a variable displacement pump 2 (hereinafter simply referred to as “pump 2”) driven by a prime mover 1 has its swash plate tilt angle controlled by a regulator 10 to adjust the discharge flow rate. . The regulator 10 includes a servo piston 21, a feedback lever 24 engaged with the servo piston 21, a compensator 30 biased in one end direction (one of the axial directions) by a control spring 31, And a compensator sleeve 32 (in other words, the compensator spool 30 is inserted). The control pressure guided to the large-diameter pressure receiving portion 22 of the servo piston 21 is adjusted by moving the compensator spool 30 and the compensator sleeve 32 in the axial direction.

And, between the compensator spool 30 and the compensator sleeve 32 of this embodiment, a pressure adjusting unit 51 for controlling the position of the servo piston 21 and a load calculating unit for horsepower control are integrally provided. Details of the pressure adjusting unit 51 will be described later.

In addition, since the double pump 2 (only one unit is shown in the figure) is used between the compensator spool 30 and the compensator sleeve 32, a plurality of discharge

pressure introduction chambers

34 and 35 are provided. Yes. In this embodiment, as the plurality of discharge

pressure introducing chambers

34 and 35, a self pressure introducing chamber 34 into which the self pressure Pd is introduced, and the direction opposite to the control spring 31 from the self pressure introducing chamber 34 (that is, the one end direction). A counterpart pressure introduction chamber 35 into which the counterpart pressure P2 is introduced is provided at a position far from the center.

Steps

42 and 43 are provided in the plurality of discharge

pressure introduction chambers

34 and 35, as will be described later. The counterpart pressure introduction chamber 35 has a smaller diameter than the self pressure introduction chamber 34.

In this embodiment, the horsepower setting pressure Pf is introduced on the side opposite to the control spring 31 of the compensator 30. By changing the horsepower setting pressure Pf, the set horsepower of the regulator 10 can be changed.

The configuration of the variable displacement pump regulator 10 will be described in more detail with reference to FIGS. The regulator casing 11 of the regulator 10 of this embodiment is fixed to the pump casing 12 with bolts 13 and 14 (FIG. 3). The pump casing 12 is provided with a servo piston 21. The regulator casing 11 is provided with a compensator spool 30 and a compensator sleeve 32 that moves in the axial direction around the compensator spool 30.

As shown in FIG. 2, the control spring 31 for urging the compensator spool 30 in one end direction is provided on the right side (one axial direction) of the compensator spool 30. As shown in FIG. The control spring 31 is displaced by moving the compensator spool 30 in the right direction by the discharge pressure introduced into the self-pressure introduction chamber 34 and the counterpart pressure introduction chamber 35 of the compensator spool 30. The control spring 31 of this embodiment is composed of two springs provided on the same axis. The position of the compensator 30 is determined by the relationship between the spring force of the control spring 31 and the three pressures (Pd, P2, Pf) acting on itself. By configuring the control spring 31 with two springs, as shown in FIG. 7 to be described later, a control line whose inclination changes midway due to a change in flow rate is approximated to an equal horsepower line. The control spring 31 may be one or more springs having a linear displacement-load characteristic or a spring having a non-linear displacement-load characteristic.

The compensator sleeve 32 is movable in the axial direction along the guide tube portion 15 provided in the regulator casing 11. One end of a feedback lever 24 is engaged with the compensator 32. The other end of the feedback lever 24 is engaged with the servo piston 21 by a control pin 27. The feedback lever 24 is provided with an engagement pin 25 at one end. The compensator sleeve 32 is provided with an engaging groove 37 on the side surface of the control spring side portion, and the engaging pin 25 of the feedback lever 24 is fitted into the engaging groove 37. The feedback lever 24 is supported on the regulator casing 11 by a support pin 26 provided at an intermediate portion, and is swung around the support pin 26 by the movement of the servo piston 21. The compensation sleeve 32 is moved in the axial direction by the oscillation of the feedback lever 24. That is, the position of the compensator 32 is determined by the position of the servo piston 21.

The plurality of discharge

pressure introducing chambers

34 and 35 described above are formed by the compensator spool 30 and the compensator sleeve 32. The compensator 30 includes a large-diameter portion 38 serving as a guide portion, a pressure-controlling land portion 33 having the same diameter as the large-diameter portion 38, and a pressure-controlling land portion 33 in order from the control spring 31. A small-diameter medium-diameter portion 39 and a small-diameter portion 40 having a diameter smaller than that of the medium-diameter portion 39 are provided, and the shaft portion 41 is integrally connected therebetween. A horsepower setting operation piston 44 is provided at the end (left direction) opposite to the control spring 31 of the compensator spool 30, and this operation piston 44 is pivoted by a cylindrical guide 16 provided on the cover 17. Guided in the direction. The compensator spool 30 and the horsepower setting operation piston 44 may be configured integrally or separately.

The compensator sleeve 32 is provided with a guide tube portion 48 for guiding the large-diameter portion 38 in the axial direction on the control spring 31 side of the inner surface formed in a cylindrical shape, and a guide tube portion 48 is guided to a predetermined position of the guide tube portion 48. A control pressure output chamber 36 having a predetermined size and a large diameter from the portion 48 is provided. The control pressure output chamber 36 is provided at a position corresponding to the pressure adjusting land portion 33 provided in the compensator 30. The control pressure output chamber 36 and the pressure adjustment land portion 33 form a pressure adjustment portion 51 that outputs a control pressure from the discharge pressure introduction chamber 34 to the large-diameter pressure receiving portion 22 of the servo piston 21. In the direction opposite to the control spring 31 of the control pressure output chamber 36, the self-pressure introduction chamber 34 having the same diameter as the guide cylinder portion 48 is formed, and in the direction opposite to the control spring 31 of the self-pressure introduction chamber 34. The counterpart pressure introduction chamber 35 is formed with an inner diameter that guides the medium diameter portion 39 in the axial direction. In the direction opposite to the control spring 31 of the counterpart pressure introducing chamber 35, the inner diameter of the compensator sleeve 32 is formed to be an inner diameter that guides the small diameter portion 40 in the axial direction.

The inner diameter portion of the self-pressure introduction chamber 34 formed between the pressure adjusting land portion 33 and the middle diameter portion 39 of the compensator spool 30 is thus formed by forming the inner surface of the compensator sleeve 32 to have different inner diameters. A step portion 42 is provided on the 39 side (opposite side of the pressure adjusting portion 51), and a step portion 43 is provided on the small diameter portion 40 side of the counterpart pressure introduction chamber 35 formed between the medium diameter portion 39 and the small diameter portion 40. It has been. The step portion 42 is formed on the inner diameter portion side of the self-pressure introduction chamber 34 due to the difference in diameter between the pressure adjusting land portion 33 and the medium diameter portion 39, and the step portion 43 is formed with the medium diameter portion 39 and the small diameter portion. It is formed on the small-diameter portion side of the counterpart pressure introducing chamber 35 by the area difference of the diameter difference from the portion 40. As described above, the diameter of the counterpart pressure introduction chamber 35 is smaller than the diameter of the self-pressure introduction chamber 34 on the inner surface of the compensator sleeve 32, and the small diameter portion 40 is axially disposed in comparison with the diameter of the counterpart pressure introduction chamber 35. There is a three-stage area difference in which the diameter of the guided portion is small.

On the other hand, between the large diameter portion 38 and the pressure adjusting land portion 36 of the compensator spool 30 is formed an oil discharge chamber 61 that guides pressure oil from a control flow path 47 (described later) to a tank passage 62 provided in the regulator casing 11. ing. Further, the compensator sleeve 32 includes an oil discharge chamber 61, a control pressure output chamber 36, a self-pressure introduction chamber 34, and a counterpart pressure introduction chamber 35, respectively, a tank passage 62, a control passage 47, an introduction passage 45 described later, and an introduction described later. Communication passages 71 to 74 (reference numerals 71 to 74 are shown only in FIG. 4) for communicating with the passage 46 are provided. Each of the communication paths 71 to 74 includes an annular groove provided on the outer peripheral surface of the compensator sleeve 32 and a plurality of through holes that penetrate the compensator sleeve 32 in the radial direction.

The self-pressure Pd introduced from the pump 2 to the small-diameter pressure receiving portion 23 of the servo piston 21 is introduced into the self-pressure introduction chamber 34 through the introduction flow path 45 and the communication path 73. A counter pressure P <b> 2 is introduced into the counter pressure introducing chamber 35 from a counter pump (not shown) through the introduction flow path 46 and the communication path 74.

Then, by these discharge pressures, the compensator 30 is moved in a direction (right direction) in which the control spring 31 is contracted against the spring force of the control spring 31. This configuration is the calculation unit 50 in the horsepower control unit of the regulator 10.

Further, when the compensator spool 30 is moved in the direction in which the control spring 31 is contracted, the pressure adjusting land portion 33 is moved in the axial direction, whereby the self-pressure introduction chamber 34 and the control pressure output chamber 36 are communicated with each other, and the self-pressure is increased. Pd is introduced as a control pressure Pcl into the large-diameter pressure receiving portion 22 of the servo piston 21 through the communication path 72 and the control flow path 47. That is, the opening area between the self-pressure introduction chamber 34 and the control pressure output chamber 36 is changed by the movement of the pressure adjusting land portion 33, and the control pressure Pcl is adjusted. This configuration is the pressure adjusting unit 51 in the horsepower control unit of the regulator 10, and the pressure adjusting unit 51 is incorporated in the calculation unit 50.

As described above, the pressure adjusting land 33 is incorporated into the calculation unit 50 of the horsepower control unit so as to form an integral configuration, whereby the calculation unit 50 causes the compensator 30 to be set in accordance with the discharge pressures of the self-pump and the counterpart pump. The mechanism for adjusting the discharge flow rate of the pump 2 by introducing the control pressure Pcl for moving the servo piston 21 by the pressure adjusting land portion 33 by moving the servo piston 21 to the position is configured compactly.

Next, the operation of the compensator spool 30 and the compensator sleeve 32 will be described with reference to FIGS. When the force from the small diameter pressure receiving portion 23 acting on the servo piston 21 and the force from the large diameter pressure receiving portion 22 are balanced, the relative positional relationship between the compensator spool 30 and the compensator sleeve 32 becomes the state shown in FIG. The chamber 36 is blocked by the pressure adjusting land portion 33 and does not communicate with either the self-pressure introduction chamber 34 or the oil discharge chamber 61.

As shown in FIG. 4, the discharge pressure Pd is introduced into the self-pressure introduction chamber 34 of the variable displacement pump regulator 10, and the total pressure of the discharge pressure (self-pressure) Pd, the counterpart pressure P2, and the set pressure Pf is controlled by the control spring. When the spring load is larger than 31, the compensator spool 30 moves toward the control spring 31. As a result, the pressure adjusting land portion 33 of the compensator spool 30 connects the self-pressure introduction chamber 34 and the control pressure output chamber 36, and the control pressure is applied from the self-pressure introduction chamber 34 to the large-diameter pressure receiving portion 22 of the servo piston 21 by this communication. Pcl is introduced. Thus, the tilt angle is controlled so that the servo piston 21 decreases the discharge flow rate of the variable displacement pump 2.

Since this embodiment is an example of a double pump, in addition to the self pressure Pd introduced into the self pressure introduction chamber 34, the compensatory spool 30 is also moved by the counterpart pressure P2 introduced into the counterpart pressure introduction chamber 35. Due to these pressures, the self-pressure flows from the self-pressure introduction chamber 34 to the control pressure output chamber 36, and the control pressure Pcl is introduced into the large-diameter pressure receiving portion 22 of the servo piston 21 through the control flow path 47. As a result, the discharge flow rate is reduced in accordance with any required horsepower of the double pump.

Then, as shown in FIG. 5, when the servo piston 21 moves to the left by the control pressure Pcl introduced into the large-diameter pressure receiving portion 22 of the servo piston 21, the feedback lever 24 swings around the support pin 26 and the compensator. The sleeve 32 is moved. When the opening between the control pressure output chamber 36 and the self pressure introducing chamber 34 is blocked by the movement of the compensator 32, the control pressure Pcl is introduced from the self pressure introducing chamber 34 into the large diameter pressure receiving portion 22. stop. In this manner, the servo piston 21 adjusts the tilt angle of the pump 2 to the tilt angle of the required discharge flow rate.

On the other hand, when the self-pressure Pd introduced into the self-pressure introduction chamber 34 or the counterpart pressure P2 introduced into the counterpart pressure introduction chamber 35 decreases from the state shown in FIG. 2, the compensator 30 moves away from the control spring 31 (leftward). The control pressure output chamber 36 communicates with the oil discharge chamber 62. As a result, the pressure oil is discharged from the large-diameter pressure receiving portion 22 to the tank passage 62 through the control passage 47, the communication passage 71, and the oil discharge chamber 61, and the servo piston 21 moves to the right and the compensator 32 moves to the left. The opening between the control pressure output chamber 36 and the oil discharge chamber 61 is closed.

In this way, the tilt angle of the pump 2 is controlled, and when the pump 2 reaches a target discharge flow rate, the position of the compensator spool 30, the compensator sleeve 32, and the servo piston 21 is maintained. As described above, the discharge flow rate control of the variable displacement pump 2 is always performed in accordance with the discharge pressure of each pump that changes depending on the work of the work machine.

FIG. 6 is a schematic diagram showing the relationship between the discharge pressure Pd and the counterpart pressure P2 acting on the compensation sleeve 32 and the force acting on the engagement pin 25 of the feedback lever 24. The compensator 32 has a self-pressure Pd force F1 introduced into the self-pressure introduction chamber 34 and a counterpart pressure P2 force F2 introduced into the counterpart pressure introduction chamber 35 due to the area difference between the

step portions

42 and 43. It acts on the

steps

42 and 43. Therefore, when the discharge pressures Pd and P2 are applied between the engagement pin 25 of the feedback lever 24 and the engagement groove 37 of the compensator sleeve 32 by these acting forces, the contact portion 28 on the control spring side is always applied. A load is acting. More specifically, when the compensator sleeve 32 is always urged to the left, a force pressed in a direction away from the control spring 31 is applied to the engagement pin 25, and the counter force is applied to the engagement groove 37. Force (force in the direction toward the control spring 31) acts.

In other words, the compensator sleeve 32 is always urged in the direction away from the control spring 31 (the anti-control spring direction) by the self-pressure Pd and the counterpart pressure P2 introduced into the self-pressure introduction chamber 34 and the counterpart pressure introduction chamber 35. A load corresponding to the resultant force F1 and F2 is always applied to these joint portions.

Therefore, when the contact portion between the engagement pin 25 of the feedback lever 24 and the engagement groove 37 which is the engagement portion is worn, the control spring side portion 28 where the engagement pin 25 and the engagement groove 37 are in contact with each other. Wear out. Further, wear due to aging progresses only on the control spring side of the engagement pin 25 and the engagement groove 37, and the displacement of the feedback lever 24 due to the positional relationship between the compensator spool 30 and the compensator sleeve 32 caused by this wear. Even if the angle changes, the servo piston 21 moves only in the direction in which the pump displacement decreases. Therefore, even if the balance position of the compensator spool 30 and the compensator sleeve 32 changes due to this wear, the discharge flow rate decreases in the horsepower control of the pump 2, that is, the horsepower control characteristic changes only in the direction of lowering the horsepower of the pump.

FIG. 7 is a diagram showing the horsepower characteristics of the variable displacement pump regulator 10 in relation to the discharge pressure and the flow rate. According to the variable displacement pump regulator 10, the engagement pin (engagement portion) 25 of the feedback lever 24 is worn against the control line 101 set to approximate the design horsepower line 100. Even if a change occurs in the control characteristic, the characteristic changes only in the direction in which the discharge flow rate decreases as in the control line 102.

That is, the control line 101 that is initially set to approximate the equal horsepower line 100 changes to the flow rate decreasing side like the control line 102 after aging, so that stable operation is possible even when used over time. The variable displacement pump regulator 10 can be configured.

As described above, according to the variable displacement pump regulator 10, the calculation unit 50 and the pressure adjustment unit 51 are integrated, and the

steps

42 and 43 are provided in the self-pressure introduction chamber 34 and the counterpart pressure introduction chamber 35. Since the compensator sleeve 32 has a difference in area where the forces F1 and F2 in the counter-control spring direction are applied by the discharge pressures Pd and P2, the discharge pressure is applied to the self-pressure introduction chamber 34 and the counterpart pressure introduction chamber 35. In the applied state, forces F1 and F2 are always applied to the compensator sleeve 32 in the anti-control spring direction.

Therefore, the engagement pin (engagement portion) 25 of the feedback lever 24 that controls the position of the compensator sleeve 32 is always in contact with the control spring side portion 28, and wear due to use over time is caused by the engagement pin 25 and the engagement groove 37. Only the control spring side portion 28 in contact with is.

In other words, the compensator sleeve 32 whose position is controlled by the feedback lever 24 by the movement of the servo piston 21 is worn even if the engagement pin 25 which is an engagement portion with the feedback lever 24 is worn over time. Is directional, and the servo piston 21 is always worn so that the feedback lever 24 tilts in the direction of decreasing the discharge flow rate of the pump 2.

Therefore, even if the contact portion between the engagement pin 25 and the engagement groove 37 is worn due to use over time, the feedback lever 24 always tilts only in the direction in which the pump tilts toward the small capacity side. Thus, the variable displacement pump regulator 10 capable of controlling the horsepower can be configured.

Further, in this embodiment, since the self-pressure Pd and the counter pressure P2 of the double pump are introduced into the calculation unit 50, the servo piston 21 is driven according to the self-pressure Pd and the counter pressure P2, and the target power is generated. The discharge flow rate of the pump 2 can be controlled so as not to exceed.

In the above embodiment, the double-pump variable displacement pump regulator 10 has been described as an example. However, the pump may have a single pump or other configuration, and is not limited to the above embodiment.

Further, in the above-described embodiment, the

step portions

42 and 43 that give the two discharge

pressure introduction chambers

34 and 35 an area difference are formed. However, the regulator 10 may have only the self-pressure introduction chamber 34. Further, there may be two or more discharge pressure introduction chambers, and a pressure necessary for flow rate control may be applied.

Furthermore, the above-described embodiment shows an example, and various modifications can be made without departing from the gist of the present invention, and the present invention is not limited to the above-described embodiment.

The variable displacement pump regulator according to the present invention can be used in hydraulic excavators, cranes, wheel loaders, bulldozers and the like used in civil engineering and construction work.

1 prime mover 2 variable displacement pump 10 regulator for variable displacement pump 11 regulator casing 21 servo piston 22 large diameter pressure receiving portion 23 small diameter pressure receiving portion 24 feedback lever 25 engagement pin (engagement portion)
30 Compensation spool 31 Control spring 32 Compensation sleeve 33 Pressure regulating land 34 Self-pressure introduction chamber (discharge pressure introduction chamber)
35 Counter pressure introduction chamber (discharge pressure introduction chamber)
36 Control pressure output chamber 37 Engagement groove 42 Step portion 43 Step portion 51 Pressure adjustment portion 101 Control line 102 Control line (hysteresis)
Pd Self pressure (Discharge pressure)
P2 Counter pressure (discharge pressure)
Pcl control pressure

Claims

A servo piston that changes the discharge flow rate of a variable displacement pump driven by a prime mover;
A feedback lever for detecting the position of the servo piston;
A compensation spool biased in one direction by a control spring;
A compensation sleeve positioned around the compensation spool and configured to move in the axial direction of the compensation spool via an engagement portion with the feedback lever by the action of the servo piston and the feedback lever; Prepared,
A discharge pressure introducing chamber into which the discharge pressure of the pump is introduced, and a pressure adjusting unit that outputs a control pressure from the discharge pressure introducing chamber to the large-diameter pressure receiving portion of the servo piston are formed by the compensator spool and the compensator sleeve. And
A regulator for a variable displacement pump, wherein the compensator is provided with a step portion in which a pressure acts in the one end direction by a discharge pressure introduced into the discharge pressure introduction chamber.
The variable capacity pump regulator according to claim 1, wherein the step portion is formed on the opposite side of the discharge pressure introducing chamber from the pressure adjusting portion by reducing the inner diameter of the compensator sleeve.
The discharge pressure introduction chamber is a self-pressure introduction chamber into which a self-pressure that is a discharge pressure of the pump is introduced, and another variable capacity is provided at a position away from the self-pressure introduction chamber on the side opposite to the pressure regulating unit. The regulator for a variable displacement pump according to claim 1 or 2, wherein a counterpart pressure introduction chamber into which a counterpart pressure as a discharge pressure of the mold pump is introduced is formed by the compensation spool and the compensation sleeve.
The counterpart pressure introduction chamber is formed with a smaller diameter than the self-pressure introduction chamber,
The regulator for a variable displacement pump according to claim 3, wherein the compensator is provided with a step portion in which a pressure acts in the one end direction by a counterpart pressure introduced into the counterpart pressure introduction chamber.