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Inventors
Watanabe, Hiroshi
Izumi, Eiki
Tanaka, Yasuo
Onoue, Hiroshi
Nakamura, Shigetaka
Application #
601798
Filed
Oct-31-1990
Published
Dec-15-1992
Current US Class
060/452
091/446
091/511
International Classes
F16D 031/02
Field of Search
60/420 60/423 60/434 60/449 60/445 60/451 60/452 91/511 91/446 417/43
Assignee
Hitachi Construction Machinery Co., Ltd. (Tokyo, JP)
Examiners
Look; Edward K.
Attorney, Agent or Firm
Fay, Sharpe, Beall, Fagan, Minnich & McKee
US Patent References
| 4617854 |
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Multiple consumer... |
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| 4809504 |
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Control system for c... |
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| 4856278 |
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Control arrangeme... |
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| 4967557 |
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Control system for l... |
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Referenced by:
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Citation
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Abstract
A control system for a hydraulic pump in a hydraulic drive circuit includes at least one hydraulic pump whose displacement volume is variable, at least one hydraulic actuator driven by a hydraulic fluid delivered from the hydraulic pump, and a flow control valve connected between the hydraulic pump and the actuator for controlling a flow rate of the hydraulic fluid supplied to the actuator. In the control system, the target value of the differential pressure between the delivery pressure of the hydraulic pump and the load pressure of the actuator is preset, and the displacement volume is varied dependent on the deviation between the differential pressure and its target value for controlling a pump delivery rate so that the differential pressure is held at the target value. The control system further influences the change rate of the delivery pressure of the hydraulic pump with respect to the change in the displacement volume of the hydraulic pump, and determines a control gain (Ki) for the change rate of the displacement volume from the received value; and controls the displacement volume of the hydraulic pump in accordance with the control gain and the differential pressure deviation.
Claims
What is claimed is:
1. A control system for a hydraulic pump in a hydraulic drive circuit comprising at least one hydraulic pump provided with displacement volume varying means, at least one hydraulic actuator driven by a hydraulic fluid delivered from said hydraulic pump, and a flow control valve connected between said hydraulic pump and said actuator for controlling a flow rate of the hydraulic fluid supplied to said actuator, wherein a target value of a differential pressure between a delivery pressure of said hydraulic pump and a load pressure of said actuator is preset, and said displacement volume varying means of said hydraulic pump is driven dependent on a deviation between said differential pressure and said target value thereof for controlling a pump delivery rate so that said differential pressure is held at said target value, said control system for a hydraulic pump further comprising:
first means for receiving at least one value; which influences a change rate of the delivery pressure of said hydraulic pump with respect to change in the displacement volume of said hydraulic pump, and determining a control gain for a change rate of the displacement volume based on the received value; and
second means for controlling said displacement volume varying means of said hydraulic pump in accordance with the control gain determined by said first means and said differential pressure deviation.
2. A control system for a hydraulic pump according to claim 1, wherein said first means determines said control gain based on said received value such that as the change rate of the delivery pressure of said hydraulic pump with respect to change in the displacement volume of said hydraulic pump becomes larger, the change rate of said displacement volume is decreased, and as the change rate of the delivery pressure of said hydraulic pump with respect to change in the displacement volume as said hydraulic pump becomes smaller, the change rate of said displacement volume is increased.
3. A control system for a hydraulic pump according to claim 1, wherein the received value of said first means is a value relating to an operated state of said flow control valve.
4. A control system for a hydraulic pump according to claim 3, wherein the value relating to an operated state of said flow control valve is the displacement volume of said hydraulic pump.
5. A control system for a hydraulic pump according to claim 3, wherein the value relating to an operated state of said flow control valve is said differential pressure deviation.
6. A control system for a hydraulic pump according to claim 3, wherein the value relating to an operated state of said flow control valve is a deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump.
7. A control system for a hydraulic pump according to claim 3, wherein the value relating to an operated state of said flow control valve includes the displacement volume of said hydraulic pump and said differential pressure deviation.
8. A control system for a hydraulic pump according to claim 3, wherein the value relating to an operated state of said flow control valve includes the displacement volume of said hydraulic pump and a deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump.
9. A control system for a hydraulic pump according to claim 1, wherein the received value of said first means is a revolution speed of said hydraulic pump.
10. A control system for a hydraulic pump according to claim 1, wherein the received value of said first means includes a value relating to an operated state of said flow control valve and a revolution speed of said hydraulic pump.
11. A control system for a hydraulic pump according to claim 4, wherein said control gain is set in a relationship that said control gain becomes larger as the displacement volume of said hydraulic pump is increased, and becomes smaller as the displacement volume is decreased.
12. A control system for a hydraulic pump according to claim 5, wherein said control gain is set in a relationship that said control gain becomes larger as said differential pressure deviation is increased, and becomes smaller as said differential pressure deviation is decreased.
13. A control system for a hydraulic pump according to claim 6, wherein said control gain is set in a relationship that said control gain becomes larger as the deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump is increased, and becomes smaller as the deviation is decreased.
14. A control system for a hydraulic pump according to claim 9, wherein said control gain is set in a relationship that said control gain becomes smaller as the revolution speed of said hydraulic pump is increased, and becomes larger as the revolution speed is decreased.
15. A control system for a hydraulic pump according to claim 1, wherein said first means includes third means for determining at least one control coefficient for arithmetic operation based on said received value, and said second means includes fourth means for determining a target displacement volume from said differential pressure deviation and said control coefficient, and controlling said displacement volume varying means of said hydraulic pump in accordance with said target displacement volume.
16. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means is the displacement volume of said hydraulic pump, and said third means calculates said control coefficient based on said displacement volume.
17. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means is said differential pressure deviation, and said third means calculates said control coefficient based on said differential pressure deviation.
18. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means is a deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump, and said third means calculates said control coefficient based on said flow rate deviation.
19. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means is a revolution speed of said hydraulic pump, and said third means calculates said control coefficient based on said revolution speed.
20. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means includes the displacement volume of said hydraulic pump and a revolution speed of said hydraulic pump, and said third means calculates said control coefficient based on these values.
21. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means includes said differential pressure deviation and a revolution speed of said hydraulic pump, and said third means calculates said control coefficient based on these values.
22. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means includes a deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump and a revolution speed of said hydraulic pump, and said third means calculates said control coefficient based on these values.
23. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means includes the displacement volume of said hydraulic pump and said differential pressure deviation, and said third means calculates said control coefficient based on these values.
24. A control system for a hydraulic pump according to claim 15, wherein the received value of said third means includes the displacement volume of said hydraulic pump and a deviation between the demanded flow rate of said flow control valve and the delivery rate of said hydraulic pump, and said third means calculates said control coefficient based on these values.
25. A control system for a hydraulic pump according to claim 20, wherein said third means calculates plural primary control coefficients dependent on said plural values, respectively, and calculates said control coefficient from said plural primary coefficients.
26. A control system for a hydraulic pump according to claim 16, wherein said control coefficient is set in a relationship that said control coefficient becomes larger as said displacement volume is increased, and becomes smaller as said displacement volume is decreased.
27. A control system for a hydraulic pump according to claim 17, wherein said control coefficient is set in a relationship that said control coefficient becomes larger as said differential pressure deviation is increased, and becomes smaller as said differential pressure deviation is decreased.
28. A control system for a hydraulic pump according to claim 18, wherein said control coefficient is set in a relationship that said control coefficient becomes larger as said flow rate deviation is increased, and becomes smaller as said flow rate deviation is decreased.
29. A control system for a hydraulic pump according to claim 19, wherein said control coefficient is set in a relationship that said control coefficient becomes smaller as said revolution speed is increased, and becomes larger as said revolution speed is decreased.
30. A control system for a hydraulic pump according to claim 16, wherein the displacement volume as said received value is a target displacement volume determined by said fourth means.
31. A control system for a hydraulic pump according to claim 16, wherein said control system further comprises means for detecting an actual displacement volume of said hydraulic pump, and the displacement volume as said received value is the detected displacement volume.
32. A control system for a hydraulic pump according to claim 17, wherein said control system further comprises means for detecting a differential pressure between the delivery pressure of said hydraulic pump and the load pressure of said actuator, and means for calculating the deviation between the detected differential pressure and preset target value of the differential pressure, and wherein the differential pressure deviation as said received value is the calculated differential pressure deviation.
33. A control system for a hydraulic pump according to claim 18, wherein said control system further comprises means for calculating a delivery rate of said hydraulic pump from the target displacement volume determined by said fourth means, and means for calculating a deviation between a demanded flow rate of said flow control valve and the detected delivery rate, and wherein the flow rate deviation as said received value is the calculated flow rate deviation.
34. A control system for a hydraulic pump according to claim 18, wherein said control system further comprises means for detecting an actual displacement volume of said hydraulic pump, means for calculating a delivery rate of said hydraulic pump from the detected displacement volume, and means for calculating a deviation between a demanded flow rate of said flow control valve and the detected delivery rate, and wherein the flow rate deviation as said received value is the calculated flow rate deviation.
35. A control system for a hydraulic pump according to claim 18, wherein said control system further comprises means for detecting an operation amount of said flow control valve, means for calculating a demanded flow rate of said flow control valve from the detected operation amount, and means for calculating a deviation between the calculated demanded flow rate and a delivery rate of said hydraulic pump, and wherein the flow rate deviation as said received value is the calculated flow rate deviation.
36. A control system for a hydraulic pump according to claim 18, wherein said hydraulic actuator and said flow control valve are each provided in plural, wherein said control system further comprises means for detecting operation amounts of said plural flow control valves, respectively, means for totaling those detected operation amounts to calculate a total demanded flow rate of said plural flow control valves, and means for calculating a deviation between the calculated demanded flow rate and a delivery rate of said hydraulic pump, and wherein the flow rate deviation as said received value is the calculated flow rate deviation.
37. A control system for a hydraulic pump according to claim 19, wherein said control system further comprises means for detecting a target revolution speed of a prime mover for driving said hydraulic pump, and the revolution speed for said hydraulic pump as said received value is the detected target revolution speed.
38. A control system for a hydraulic pump according to claim 19, wherein said control system further comprises means for detecting an actual revolution speed of a prime mover for driving said hydraulic pump, and the revolution speed of said hydraulic pump as said received value is the detected revolution speed.
39. A control system for a hydraulic pump according to claim 19, wherein said control system further comprises means for detecting an actual revolution speed of said hydraulic pump, and the revolution speed of said hydraulic pump as said received value is the detected revolution speed.
40. A control system for a hydraulic pump according to claim 15, wherein said third means includes means for presetting a basic value of said control coefficient, means for calculating a modifying coefficient of said basic value dependent on said received value, and means for multiplying said basic value by said modifying coefficient to calculate said control coefficient.
41. A control system for a hydraulic pump according to claim 15, wherein said fourth means includes means for multiplying said differential pressure deviation by said control coefficient to calculate a target change rate of said displacement volume, and means for adding said target change rate to a target displacement volume determined by calculation in the last cycle to determine said target displacement volume.
42. A control system for a hydraulic pump according to claim 15, wherein said fourth means includes means for multiplying said differential pressure deviation by said control coefficient to calculate said target displacement volume.
43. A control system for a hydraulic pump according to claim 15, wherein said third means includes means for calculating, as said control coefficient, a first control coefficient for integral control, and means for calculating a second control coefficient for proportional compensation, and said fourth means includes means for calculating a target displacement volume for integral control from said differential pressure deviation and said first control coefficient, means for calculating a modification value for proportional compensation from said differential pressure deviation and said second control coefficient, and means for calculating said target displacement volume from said target displacement volume for the integral control and said modification value for the proportional compensation.
Description
TECHNICAL FIELD
The present invention relates to a control system for a hydraulic pump in a hydraulic drive circuit for use in hydraulic machines such as hydraulic excavators and cranes, and more particularly to a control system for a hydraulic pump in a hydraulic drive circuit of load sensing control type which controls a pump delivery rate in such a manner as to hold the delivery pressure of the hydraulic pump higher than the load pressure of a hydraulic actuator, by a fixed value.
BACKGROUND ART
Hydraulic drive circuits for use in hydraulic machines such as hydraulic excavators and cranes each include at least one hydraulic pump, at least one hydraulic actuator driven by a hydraulic fluid delivered from the hydraulic pump, and a flow control valve connected between the hydraulic pump and the actuator for controlling a flow rate of the hydraulic fluid supplied to the actuator. It is known that some of those hydraulic drive circuits employs a technique called load sensing control (LS control) for controlling the delivery rate of the hydraulic pump. The load sensing control is to control the delivery rate of the hydraulic pump such that a delivery pressure of the hydraulic pump is held at a fixed value higher than a load pressure of the hydraulic actuator. This causes the delivery rate of the hydraulic pump to be controlled dependent on the load pressure of the hydraulic actuator, and hence permits economic operation.
Meanwhile, the load sensing control is carried out by detecting a differential pressure (LS pressure) between the delivery pressure and the load pressure, and controlling the displacement volume of the hydraulic pump, or the position (tilting amount) of a swash plate in the case of a swash plate pump, in response to a deviation between the LS differential pressure and a differential pressure target value. Conventionally, the detection of the differential pressure and the control of tilting amount of the swash plate have usually been carried out in a hydraulic manner as disclosed in JP, A, 60-11706, for example. This conventional arrangement will briefly be described below.
A pump control system disclosed in JP, A, 60-11706 comprises a control valve having one end subjected to the delivery pressure of a hydraulic pump and the other end subjected to both the maximum load pressure among a plurality of actuators and the urging force of a spring, and a cylinder unit operation of which is controlled by a hydraulic fluid passing through the control valve for regulating the swash plate position of the hydraulic pump. The spring at one end of the control valve is to set a target value of the LS differential pressure. Depending on the deviation occurred between the LS differential pressure and the target value, the control valve is driven and the cylinder unit is operated to regulate the swash plate position, whereby the pump delivery rate is controlled so that the LS differential pressure is held at the target value. The cylinder unit has a spring built therein to apply an urging force in opposite relation to the direction in which the cylinder unit is driven upon inflow of the hydraulic fluid.
However, the above conventional control system for the hydraulic pump has had problems below.
In the conventional pump control system, the tilting speed of a swash plate of the hydraulic pump is determined dependent on the flow rate of the hydraulic fluid flowing into the cylinder unit, while the flow rate of the hydraulic fluid is determined dependent on both an opening, i.e., a position, of the control valve and setting of the spring in the cylinder unit and, in turn, the position of the control valve is determined by the relationship between the urging force of the LS differential pressure and the spring force for setting the target value. Here, the spring of the control valve and the spring of the cylinder unit each have a fixed spring constant. Accordingly, a control gain for the tilting speed of the swash plate dependent on the deviation between the LS differential pressure and the target value thereof is always constant. The control gain, i.e., the spring constants of the two springs, are set in such a range that change in the pump delivery pressure will not cause hunting and the pump is kept from coming into disablement of control on account of change in the delivery rate upon change in the swash plate position.
In the LS control, the delivery pressure of the hydraulic pump is determined dependent on a difference between the flow rate of the hydraulic fluid flowing into a line, extending from the hydraulic pump to the flow control valve, and the flow rate of the hydraulic fluid flowing out of the line, as well as a volume into which the delivered hydraulic fluid is allowed to flow. Therefore, when the operation (input) amount of the flow control valve (i.e., the demanded flow rate) is small, the opening of the flow control valve is so reduced that the small line volume between the hydraulic pump and the flow control valve plays a predominant factor. As a result, the delivery pressure is largely varied even with slight change in the flow rate upon change in the swash plate position. On the other hand, when the operation amount of the flow control valve is increased to enlarge the opening thereof, the large line volume between the pump and an actuator now takes part in pressure change, whereby change in the delivery pressure upon change in the delivery rate is reduced.
Accordingly, in order to prevent the occurrence of hunting over a range of the entire operation amount (opening) of the flow control valve, the above-mentioned control gain, i.e., the spring constants of the two springs, are set to provide such a tilting speed of the swash plate as to prevent the pressure change from hunting at the small opening of the flow control valve for the positive LS control.
With the control gain set as explained above, under a condition that the operation amount of the flow control valve is small and hence its opening is small, i.e., when the hydraulic pump is at the low delivery rate, change in the delivery rate produce proper change in pressure and will not cause hunting. But under a condition that the operation amount of the flow control valve is large and hence its opening is large, i.e., when the hydraulic pump is at the high delivery rate, the tilting speed of the swash plate dependent on change in the delivery rate is restricted by the above-mentioned control gain, and too small pressure change makes it difficult to control the delivery pressure with a good response. For instance, therefore, when an operating lever of the flow control valve is operated in a large stroke to increase the opening of the flow control valve, an operator is forced to feel that the actuator is too slow in action.
Further, when the operating lever is operated at small speeds and hence the deviation between the demanded flow rate of the flow control valve and the delivery rate of the hydraulic pump is small, the deviation between the LS differential pressure and the differential pressure target value is also small, and thus the change in pressure upon change in the tilting speed of the swash plate, i.e., the change in the delivery rate is sufficient to realize demanded speed change of the actuator. On the contrary, when the operating lever of the flow control valve is operated at large speeds to abruptly increase the opening of the flow control valve, there occurs a large difference between the demanded flow rate of the flow control valve and the delivery rate of the hydraulic pump, which also increases the deviation between the LS differential pressure and the differential pressure target value. Under this condition, the tilting speed of the swash plate is restricted by the above-mentioned control gain, and hence it takes a time for the once reduced differential pressure to return to its target value. As a consequence, the demanded speed change of the actuator cannot be realized, causing the operator to feel that the actuator is too slow in action.
The above description has been made without taking into account a revolution speed of the hydraulic pump. The delivery rate of the hydraulic pump is also influenced by the pump revolution speed such that when the pump revolution speed is high, even slight change in the swash plate position produce large flow rate change and hence large pressure change. In construction machines such as hydraulic excavators, a hydraulic pump is driven by a prime mover via a speed reducer and, as a revolution speed of the prime mover changes, a pump revolution speed is also changed. It is hence required that change in the flow rate dependent on change in the swash plate position be kept within a proper range even at the maximum pump revolution speed, in order to prevent the occurrence of hunting over an entire range of the pump revolution speed, i.e., the revolution speed of the prime mover, and to ensure the positive LS control. For this purpose, the above-mentioned control gain, i.e., the spring constants of the two springs, are also so set as to prevent the pressure change from hunting at the maximum pump revolution speed (or the revolution speed of the prime mover).
With the control gain thus set, when the revolution speed of the hydraulic pump is at maximum, change in the swash plate position produces satisfactory change in the delivery rate to realize the demanded speed change of the actuator. However, when the pump revolution speed is low, the tilting speed of the swash plate is restricted by the above-mentioned control gain, and change in the swash plate position produces small change in the delivery rate. Consequently, the demanded speed change of the actuator cannot be realized and the operator is forced to feel that the actuator is too slow in action.
An object of the present invention is to provide a control system for a hydraulic pump which permits, in a hydraulic drive circuit of load sensing control type, to properly control a change rate of the delivery rate with respect to change in the displacement volume of the hydraulic pump to prevent the occurrence of hunting due to an abrupt change of the pump delivery pressure and achieve a prompt response.
SUMMARY
To achieve the above object, according to the present invention, there is provided a control system for a hydraulic pump in a hydraulic drive circuit comprising at least one hydraulic pump provided with displacement volume varying means, at least one hydraulic actuator driven by a hydraulic fluid delivered from said hydraulic pump, and a flow control valve connected between said hydraulic pump and said actuator for controlling a flow rate of the hydraulic fluid supplied to said actuator, wherein a target value of a differential pressure between a delivery pressure of said hydraulic pump and a load pressure of said actuator is preset, and said displacement volume varying means of said hydraulic pump is driven dependent on a deviation between said differential pressure and said target value thereof for controlling a pump delivery rate so that said differential pressure is held at said target value, said control system for a hydraulic pump further comprising first means for receiving at least one value which influences a change rate of the delivery pressure of said hydraulic pump with respect to change in the displacement volume of said hydraulic pump, and determining a control gain for a change rate of the displacement volume based on the received value; and second means for controlling said displacement volume varying means of said hydraulic pump in accordance with the control gain determined by said first means and said differential pressure deviation.
Thus, a value of at least one parameter is entered which influences a change rate of the delivery pressure of the hydraulic pump with respect to change in the displacement volume of the hydraulic pump, and the control gain for the change rate of the displacement volume is determined based on the entered value to control the varying speed of the displacement volume. The change rate of the delivery rate with respect to change in the displacement volume of the hydraulic pump is thereby controlled properly to permit a prompt response without making the pump delivery pressure so abruptly changed as to cause hunting.
The first means preferably determines the control gain based on the aforesaid received value such that as the change rate of the delivery pressure of the hydraulic pump with respect to change in the displacement volume of the hydraulic pump becomes larger, the change rate of the displacement volume is decreased, and as the change rate of the delivery pressure of the hydraulic pump with respect to change in the displacement volume of the hydraulic pump becomes smaller, the change rate of the displacement volume is increased.
Preferably, the first means includes third means for determining at least one control coefficient for arithmetic operation based on the aforesaid received value, and the second means includes fourth means for determining a target displacement volume from the differential pressure deviation and the control coefficient, and controlling the displacement volume varying means of the hydraulic pump in accordance with the target displacement volume.
The received value of the third means is prefereably the displacement volume of the hydraulic pump, and the third means calculates the control coefficient based on the displacement volume.
Further, the received value(s) of the third means may be the differential pressure deviation; a deviation between a demanded flow rate of the flow control valve and the delivery rate of the hydraulic pump; a revolution speed of the hydraulic pump; the displacement volume of the hydraulic pump and the revolution speed of the hydraulic pump; the differential pressure deviation and the revolution speed of the hydraulic pump; the flow rate deviation and the revolution speed of the hydraulic pump; the displacement volume of the hydraulic pump and the differential pressure deviation; or the displacement volume of the hydraulic pump and the flow rate deviation.
When the receiving the plurality of values, the third means calculates a plurality of primary control coefficients dependent on the received values, respectively, and then calculates the control coefficient from the plurality of primary control coefficients.
In the case where the aforesaid received value is the displacement volume of the hydraulic pump, the control coefficient is set in a relationship that it becomes larger as the displacement volume is increased, and becomes smaller as the displacement volume is decreased.
In the case where the aforesaid received value is the differential pressure deviation, the control coefficient is set in a relationship that it becomes larger as the differential pressure deviation is increased, and becomes smaller as the differential pressure deviation is decreased.
In the case where the aforesaid received value is the flow rate deviation, the control coefficient is set in a relationship that it becomes larger as the flow rate deviation is increased, and becomes smaller as the flow rate deviation is decreased.
In the case where the aforesaid received value is the revolution number of the hydraulic pump, the control conefficient is set in a relationship that it becomes smaller as the revolution speed is increased, and becomes larger as the revolution speed is decreased.
The displacement volume as the aforesaid received value may be a target displacement volume determined by the fourth means. Further, the control system of the present invention may further comprise means for detecting an actual displacement volume of the hydraulic pump, and the displacement volume as the aforesaid received value may be the detected displacement volume.
The control system of the present invention may further comprise means for detecting a differential pressure between the delivery pressure of the hydraulic pump and the load pressure of the actuator, and means for calculating a deviation between the detected differential pressure and a preset target value of the differential pressure, and the differential pressure deviation as the aforesaid received value may be this calculated differential pressure deviation.
The control system of the present invention may further comprise means for calculating a delivery rate of the hydraulic pump from the target displacement volume determined by the fourth means, and means for calculating a deviation between the demanded flow rate of the flow control valve and the detected delivery rate, and the flow rate deviation as the aforesaid received value may be this calculated flow rate deviation.
The control system of the present invention may further comprise means for detecting the actual displacement volume of the hydraulic pump, means for calculating the delivery rate of the hydraulic pump from the detected displacement volume, and means for calculating a deviation between the demand flow rate of the flow control valve and the detected delivery rate, and the flow rate deviation as the aforesaid received value may be this calculated flow rate deviation.
The control system of the present invention may further comprise means for detecting an operation amount of the flow control valve, means for calculating the demanded flow rate of the flow control valve from the detected operation amount, and means for calculating a deviation between the calculated demanded flow rate and the delivery rate of the hydraulic pump, and the flow rate deviation as the aforesaid received value may be this calcultated flow rate deviation.
In the case where the hydraulic actuator and the flow control valve are each provided in plural, the control system of the present invention may further comprise means for detecting operation amounts of the plural flow control valves, respectively, means for totaling those detected operation amounts to calculate a total demanded flow rate of the plural flow control valves, and means for calculating a deviation between the calculated demanded flow rate and the delivery rate of the hydraulic pump, and the flow rate deviation as the aforesaid received value may be this calculated flow rate deviation.
The control system of the present invention may further comprise means for detecting a target revolution speed of a prime mover for driving the hydraulic pump, and the revolution speed of the hydraulic pump as the aforesaid received value is this detected target revolution speed.
The control system of the present invention may further comprise means for detecting an actual revolution speed of the prime mover for driving the hydraulic pump, and the revolution speed of the hydraulic pump as the aforesaid received value is this detected revolution speed.
The control system of the present invention may further comprise means for detecting an actual revolution speed of the hydraulic pump, and the revolution speed of the hydraulic pump as the aforesaid received value is this detected revolution speed.
Preferably, the third means includes means for presettting a basic value of the control coefficient, means for calculating a modifying coefficient of the basic value dependent on the aforesaid received value, and means for multiplying the basic value by the modifying coefficient to calculate the control coefficient.
Preferably, the fourth means includes means for multiplying the differential pressure deviation by the control coefficient to calculate a target change rate of the displacement volume, and means for adding the target change rate to a target displacement volume determined by calculation in the last cycle to determine the target displacement volume.
The fourth means may includes means for multiplying the differential pressure deviation by the control coefficient to calculate the target displacement volume. Further, the third means may include means for calculating, as the control coefficient, a first control coefficient for integral control, and means for calculating a second control coefficient for proportional compensation, and the fourth means may include means for calculating a target displacement volume for the integral control from the differential pressure deviation and the first control coefficient, means for calculating a modification value for proportional compensation from the differential pressure deviation and the second control coefficient, and means for calculating the target displacement volume from the target displacement volume for the integral control and the modification value for the proportional compensation.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic diagram of a hydraulic drive circuit of load sensing control type equipped with a control system for a hydraulic pump according to a first embodiment of the present invention;
FIG. 2 is a schematic diagram showing arrangement of a swash plate position controller;
FIG. 3 is a schematic diagram showing arrangement of a control unit;
FIG. 4 is a flowchart showing the control sequence carried out in the control unit;
FIG. 5 is a flowchart showing details of a step of calculating a control coefficient Ki in the flowchart shown in FIG. 4;
FIG. 6 is a characteristic graph showing the relationship between a swash plate position and a modifying coefficient Kr;
FIG. 7 is a flowchart showing details of a step of calculating a swash plate target position of a hydraulic pump in the flowchart of FIG. 4;
FIG. 8 is a flowchart showing details of a step of controlling the swash plate position of the hydraulic pump in the flowchart shown of FIG. 4;
FIG. 9 is a block diagram showing control steps of the first embodiment together in the form of blocks;
FIG. 10 is a chart showing change in the opening of a flow control valve, the LS differential pressure, the control coefficient and the swash plate position over time, for explaining operation of the first embodiment;
FIG. 11 is a block diagram similar to FIG. 9, showing a modification of the first embodiment;
FIG. 12 is a block diagram similar to FIG. 9, showing a control system for a hydraulic pump according to a second embodiment of the present invention;
FIG. 13 is a block diagram similar to FIG. 9, showing a control system for a hydraulic pump according to a third embodiment of the present invention;
FIG. 14 is a flowchart showing the control sequence for a control system for a hydraulic pump according to a fourth embodiment of the present invention;
FIG. 15 is a flowchart showing details of a step of calculating a control coefficient Ki in the flowchart shown in FIG. 14;
FIGS. 16(a)-16(d) are characteristic views each showing the relationship between a differential pressure deviation .DELTA. (.DELTA.P) and a modifying coefficient Kr;
FIG. 17 is a flowchart showing details of a step of calculating a swash plate target position of the hydraulic pump in the flowchart of FIG. 14;
FIG. 18 is a block diagram showing control steps of the fourth embodiment together in the form of blocks;
FIG. 19 is a chart showing change in the opening of a flow control valve, the LS differential pressure, the control coefficient and the swash plate position over time, for explaining operation of the fourth embodiment;
FIGS. 20 and 21 are block diagrams similar to FIG. 18, each showing a modification of the fourth embodiment;
FIG. 22 is a schematic diagram of a hydraulic drive circuit of load sensing control type equipped with a control system for a hydraulic pump according to a fifth embodiment of the present invention;
FIG. 23 is a flowchart showing the control sequence in the fifth embodiment;
FIG. 24 is a flowchart showing details of a step of calculating a control coefficient Ki in the flowchart shown in FIG. 23;
FIG. 25 is a characteristic graph showing the relationship between a flow rate deviation .DELTA.X and a modifying coefficient Kr;
FIG. 26 is a block diagram showing control steps of the fifth embodiment together in the form of blocks;
FIG. 27 is a chart showing change in the opening of a flow control valve, the LS differential pressure, the control coefficient and the swash plate position over time, for explaining operation of the fifth embodiment;
FIGS. 28-30 are block diagrams similar to FIG. 26, each showing a modification of the fifth embodiment;
FIG. 31 is a schematic diagram of a hydraulic drive circuit of load sensing control type equipped with a control system for a hydraulic pump according to a sixth embodiment of the present invention;
FIG. 32 is a flowchart showing the control sequence in the sixth embodiment;
FIG. 33 is a flowchart showing details of a step of calculating a control coefficient Ki in the flowchart shown in FIG. 32;
FIG. 34 is a characteristic graph showing the relationship between a target revolution speed Nr and a modifying coefficient Kr;
FIG. 35 is a block diagram showing control steps of the sixth embodiment together in the form of blocks;
FIGS. 36 and 37 are each a chart showing change in the opening of a flow control valve, the target revolution speed, the control coefficient, the LS differential pressure, the swash plate position and the pump delivery rate over time, for explaining operation of the sixth embodiment;
FIG. 38 is a block diagram of a control system for a hydraulic pump according to a seventh embodiment of the present invention;
FIG. 39 is a block diagram showing a control system for the hydraulic pump according to a modification of the seventh embodiment;
FIG. 40 is a block diagram of a control system for a hydraulic pump according to an eighth embodiment of the present invention; and
FIGS. 41 and 42 are each a block diagram showing a control system for the hydraulic pump according to a modification of the eighth embodiment.
DETAILED DESCRIPTION
Hereinafter, several embodiments of the present invention will be described with reference to the accompanying drawings.
FIRST EMBODIMENT
To begin with, a first embodiment of the present invention will be explained by referring to FIGS. 1-10.
In FIG. 1, a hydraulic drive circuit according to this embodiment comprises a hydraulic pump 1, a plurality of hydraulic actuators 2, 2A driven by a hydraulic fluid delivered from the hydraulic pump 1, flow control valves 3, 3A connected between the hydraulic pump 1 and the actuators 2, 2A for controlling flow rates of the hydraulic fluid supplied to the actuators 2, 2A dependent on operation of operating levers 3a, 3b, respectively, and pressure compensating valves 4, 4A for holding constant differential pressures between the upstream and downstream sides of the flow control valves 3, 3A, i.e., differential pressures across the valves, to control the flow rates of the hydraulic fluid passing through the flow control valves 3, 3A to values in proportion to openings of the flow control valves 3, 3A, respectively. A set of the flow control valve 3 and the pressure compensating valve 4 constitutes one pressure compensated flow control valve, while a set of the flow control valve 3A and the pressure compensating valve 4A constitutes another pressure compensated flow control valve. The hydraulic pump 1 has a swash plate 1a as a displacement volume varying mechanism.
The hydraulic pump 1 is controlled in its delivery rate by a control system of this embodiment which comprises a differential pressure sensor 5, a swash plate position sensor 6, a control unit 7 and a swash plate position controller 8. The differential pressure sensor 5 detects a differential pressure between a load pressure of the actuator 2 or 2A on the higher side selected by a shuttle valve 9, i.e., a maximum load pressure PL, and a delivery pressure Pd of the hydraulic pump 1 (i.e., an LS differential pressure), and converts it to an electric signal .DELTA.P for outputting to the control unit 7. The swash plate position sensor 6 detects a position (tilting amount) of a swash plate 1a of the hydraulic pump 1 and converts it to an electric signal .theta. for outputting to the control unit 7. The control unit 7 calculates a drive signal for the swash plate 1a of the hydraulic pump 1 based on the electric signals .DELTA.P, .theta., and outputs the drive signal to swash plate position controller 8. In response to the drive signal from the control unit 7, the swash plate position controller 8 drives the swash plate 1a for controlling the pump delivery rate.
The swash plate position controller 8 is constituted as a hydraulic drive device of electro-hydraulic servo type, for example, as shown in FIG. 2.
More specifically, the swash plate position controller 8 has a servo piston 8b for driving the swash plate 1a of the hydraulic pump 1, the servo piston 8b being housed in a servo cylinder 8c. A cylinder chamber of the servo cylinder 8c is partitioned by the servo piston 8b into a left-hand chamber 8d and a right-hand chamber 8e. These chambers are formed such that the cross-sectional area D of the left-hand chamber 8d is larger than the cross-sectional area d of the right-hand chamber 8e.
The left-hand chamber 8d of the servo cylinder 8c is communicated with a hydraulic source 10 such as a pilot pump via a line 8f, and the right-hand chamber 8e of the servo cylinder 8c is communicated with the hydraulic source 10 via a line 8i, the line 8f being communicated with being communicated with a reservoir (tank) 11 via a return line 8j. A solenoid valve 8g is interposed in the line 8f, and a solenoid valve 8h is interposed in the return line 8j. These solenoid valves 8g, 8h are each a normally closed solenoid valve (with the function of returning to a closed state upon de-energization), and switched over by the drive signal from the control unit 7.
When the solenoid valve 8g is energized (turned on) for switching to its open position B, the left-hand chamber 8d of the servo cylinder 8c is communicated with the hydraulic source 10, whereupon the servo piston 8b is forced to move rightwardly on the drawing due to the difference in the cross-sectional area between the left-hand chamber 8d and the right-hand chamber 8e. This increases a tilting angle of the swash plate 1a of the hydraulic pump 1 and hence the delivery rate. When the solenoid valve 8g and the solenoid valve 8h are both de-energized (turned off) for returning to their closed positions A, the oil passage leading to the left-hand chamber 8d is cut off and the servo piston 8b remains rest at the then position. The tilting angle of the swash plate 1a of the hydraulic pump 1 is thereby kept constant, and so is the delivery rate. When the solenoid valve 8h is energized (turned on) for switching to its open position B, the left-hand chamber 8d of the servo cylinder 8c is communicated with the reservoir 11 to reduce the pressure in the left-hand chamber 8d, whereby the servo piston 8b is forced to move leftwardly on the drawing with the pressure in the right-hand chamber 8e. This decreases the tilting angle of the swash plate 1a of the hydraulic pump 1 and hence the delivery rate.
The control unit 7 is constituted by a microcomputer and, as shown in FIG. 3, comprises an A/D converter 7a for converting the differential pressure signal .DELTA.P outputted from the differential pressure sensor 5 and the swash plate position signal .theta. outputted from the swash plate position sensor 6 to digital signals, a central processing unit (CPU) 7b, a read only memory (ROM) 7c for storing a program for the control sequence, a random access memory (RAM) 7d for temporarily storing numerical values under calculations, an I/O interface 7e for outputting the drive signals, and amplifiers 7g, 7h connected to the aforesaid solenoid valves 8g, 8h, respectively.
The control unit 7 calculates a swash plate target position .theta.o from the differential pressure signal .DELTA.P outputted from the differential pressure sensor 5 based on the program for the control sequence stored in the ROM 7c, and creates the drive signals from the swash plate target position .theta.o and the swash plate position signal .theta. outputted from the swash plate position sensor 6 for making a deviation therebetween zero, followed by outputting the drive signals to the solenoid valves 8g, 8h of the swash plate position controller 8 from the amplifiers 7g, 7h via the I/O interface 7e. The swash plate 1a of the hydraulic pump 1 is thereby controlled so that the swash plate position signal .theta. coincides with the swash plate target position .theta..
Function and operation of this embodiment will be described below in detail by referring to a flowchart, shown in FIG. 4, of a program for the control sequence stored in the ROM 7c.
First, in a step 100, respective outputs of the differential pressure sensor 5 and the swash plate position sensor 6 are entered to the control unit via the A/D converter 7a and stored in the RAM 7d as the differential pressure signal .DELTA.P and the swash plate position signal .theta..
Then, in a step 110, the control unit calculates a control coefficient Ki used for controlling a tilting speed of the swash plate 1a. FIG. 5 shows details of the step 110. In a step 111 of FIG. 5, a modifying coefficient Kr is calculated from the swash plate target position .theta.o-1 which has been calculated in the last cycle. The calculation is made by previously storing table data as shown in FIG. 6 in the ROM 7c, and reading the modifying coefficient Kr corresponding to the swash plate target position .theta.o-1 from the table data. Here, the relationship of .theta.o-1 versus Kr shown in FIG. 6 is set such that when the swash plate target position is small, the control coefficient Ki determined in a step 112 described later takes a small value which enables to perform stable control without making the delivery pressure of the hydraulic pump 1 so abruptly changed as to cause hunting, and when the swash plate target position is large, it takes a sufficient value to provide a prompt response by avoiding slow change in the delivery pressure. Notice that instead of storing the modifying coefficient Kr in the form of table data, the modifying coefficient Kr may be determined through arithmetic operations by programming the calculation formula in advance.
Then, in a step 112, the modifying coefficient Kr is multiplied by a preset basis value Kio of the control coefficient to obtain the control coefficient Ki. In this case, the basic value Kio of the control coefficient is given by a value which is optimum when the swash plate target position takes a maximum value (.theta.omax). The modifying coefficient Kr is therefore set such that, as shown in FIG. 6, it becomes 1 when the swash plate target position is at maximum (.theta.omax), and it takes a smaller value (<1) as the swash plate target position is decreased. Alternatively, the basic value Kio may be given by a value which is optimum when the swash plate target position takes a minimum value. In this case, the modifying coefficient Kr may be set such that it becomes 1 when the swash plate target position is at minimum, and it takes a larger value (>1) as the swash plate target position is increased. As a further alternative, the basic value Kio may be given by a value which is optimum when the swash plate target position is intermediate between maximum and minimum. In this case, the modifying coefficient Kr may be set such that it becomes larger (>1) as the swash plate target position is increased from the intermediate, and it becomes smaller (>1) as the swash plate target position is decreased. In either case, the control coefficient Ki is obtained as the same value.
Next, returning to FIG. 4, a step 120 calculates a swash plate target position (i.e., a target tilting amount) of the hydraulic pump through integral control. FIG. 7 shows details of the step 120. In a step 121 of FIG. 7, a deviation .DELTA. (.DELTA.P) between a present target value .DELTA.Po of the differential pressure and the differential pressure signal .DELTA.P entered in the step 100 is calculated.
Then, in a step 122, an increment .DELTA..theta..sub..DELTA.P of the swash plate target position is calculated. Specifically, the control coefficient Ki determined in the step 110 is multiplied by the above differential pressure deviation .DELTA. (.DELTA.P) to obtain the increment .DELTA..theta..sub..DELTA.P of the swash plate target position.
Assuming that a period of time required for the program proceeding from the step 100 to 130 (i.e., cycle time) is tc, the increment .DELTA..theta..sub..DELTA.P of the swash plate target position represents an increment of the swash plate target position for the cycle time tc and hence .DELTA..theta..sub..DELTA.P /tc gives a target tilting speed of the swash plate.
Then, in a step 123, the increment .DELTA..theta..sub..DELTA.P is added to the swash plate target position .theta.o-1 which has been calculated in the last cycle, to obtain the current (new) swash plate target position .theta.o.
Next, returning to FIG. 4, a step 130 controls the tilting position (tilting amount) of the hydraulic pump. FIG. 8 shows details of the control. In a step 131 of FIG. 8, a deviation Z between the swash plate target position .theta.o calculated in the step 120 and the swash plate position signal .theta. entered in the step 100 is calculated.
Then, in a step 132, it is determined whether an absolute value of the deviation Z is within a dead zone .DELTA. for the swash plate position control. If .vertline.Z.vertline. is determined to be smaller than the dead zone .DELTA. (.vertline.Z.vertline.<.DELTA.), the control flow proceeds to a step 134 where OFF signals are outputted to the solenoid valves 8g, 8h for rendering the swash plate position fixed. If .vertline.Z.vertline. is determined to be not smaller than the dead zone .DELTA. (.vertline.Z.vertline..gtoreq..DELTA.) in the step 132, the control flow proceeds to a step 133. The step 133 determines whether Z is positive or negative. If Z is determined to be positive (Z>0), the control flow proceeds to step 135. In the step 135, an ON and OFF signal are outputted to the solenoid valves 8g and 8h, respectively, for moving the swash plate position in the direction to increase.
If Z is determined to be zero or negative (Z.ltoreq.0) in the step 133, the control flow proceeds to step 136. In the step 136, an OFF and ON signal are outputted to the solenoid valves 8g and 8h, respectively, for moving the swash plate position in the direction to decrease.
Through the foregoing steps 131-136, the swash plate position is so controlled as to coincide with the target position. Also, the above steps 110-130 are carried out once for the cycle time tc mentioned above, resulting in that the tilting speed of the swash plate 1a is controlled to the aforesaid target speed .DELTA..theta..sub..DELTA.P /tc.
The above-explained control steps are shown together in FIG. 9 at 200 in the form of blocks. In FIG. 9, blocks 202-204 correspond to the step 110, blocks 201, 205, 206 correspond to the step 120, and blocks 207-209 correspond to the step 130.
Operation of this embodiment thus arranged will be described below by mainly referring to FIGS. 1 and 9.
In FIG. 1, when the operating lever 3a of the actuator 2, for example, is operated to open the flow control valve 3 to an arbitrary degree of opening, the delivery pressure of the hydraulic pump 1 is lowered to reduce the differential pressure between the pump delivery pressure Pd and the load pressure PL of the actuator 2, i.e., the LS differential pressure .DELTA.P is detected by the differential pressure sensor 5. For controlling the LS differential pressure .DELTA.P to a predetermined value, the deviation .DELTA. (.DELTA.P) between the detected differential pressure .DELTA.P and the differential pressure target value .DELTA.Po preset in the control unit 7 is first calculated. Then, this differential pressure deviation .DELTA. (.DELTA.P) is multiplied by the control coefficient Ki to determine the increment of the swash plate target position (tilting amount), i.e., the target tilting speed .DELTA..theta..sub..DELTA.P of the swash plate. This increment is added to the swash plate target value .theta.o-1 in the last cycle to calculate the new swash plate target position .theta.o. The swash plate is driven at the tilting speed of .DELTA..theta..sub..DELTA.P so as to make the actual swash plate position coincident with the swash plate target position .theta.o, thereby controlling the LS differential pressure .DELTA.P. As a result, the delivery rate of the hydraulic pump 1 is controlled so that the LS differential pressure .DELTA.P is held at the target value .DELTA.Po.
Now, when the tilting amount of the swash plate 1a is small and hence the swash plate target position .theta.o is small, the modifying coefficient Kr calculated in the block 202 of FIG. 2 also takes a small value (<1), and so does the control coefficient Ki obtained by multiplying the modifying coefficient Kr by the basic value Kio. Consequently, the swash plate target tilting speed .DELTA..theta..sub..DELTA.P is calculated as a small value, and the swash plate 1a is driven at the resultant small tilting speed.
Further, when the tilting amount of the swash plate 1a is large and hence the swash plate target position .theta.o is large, the modifying coefficient Kr calculated in the block 202 of FIG. 2 also takes a large value (.apprxeq.1), and so does the control coefficient Ki. Consequently, the swash plate target tilting speed .DELTA..theta..sub..DELTA.P is calculated as a large value, and the swash plate 1a is driven at the resultant large tilting speed.
Meanwhile, in the foregoing LS control, the delivery pressure of the hydraulic pump 1 is determined dependent on a difference between the flow rate of the hydraulic fluid flowing into a line, extending from the hydraulic pump 1 to the flow control valve 3, and the flow rate of the hydraulic fluid flowing out of the line, as well as a volume into which the delivered hydraulic fluid is allowed to flow. Therefore, when the opening of the flow control valve 3 is small, the line is so restricted by the flow control valve 3 that the small line volume between the hydraulic pump 1 and the flow control valve 3 plays a predominant factor. As a result, the delivery pressure is largely varied even with slight change in the flow rate upon change in the swash plate position. On the other hand, when the opening of the flow control valve 3 is large, the line is less restricted by the flow control valve 3 and the large line volume between the pump 1 and the actuator 2 now takes part in pressure change, whereby change in the delivery pressure upon change in the delivery rate is reduced. Stated otherwise, when the opening of the flow control valve 3 is small, the control system is in a condition likely to cause hunting, and when the opening thereof is large, it is in a condition difficult to control the delivery pressure promptly in response to change in the delivery rate.
With this embodiment, as described above, when the opening of the flow control valve 3 is small, the swash plate target tilting speed .DELTA..theta..sub..DELTA.P is calculated as a small value, and the tilting speed of the swash plate 1a becomes small. It is therefore possible to perform stable control without making the delivery pressure so abruptly changed as to cause hunting.
Also, when the opening of the flow control valve 3 is large, the swash plate target tilting speed .DELTA..theta..sub..DELTA.P is calculated as a large value, and the tilting speed of the swash plate 1a becomes large. It is therefore possible to perform stable control with a good response, while avoiding too slow change in the delivery pressure.
For instance, when the operating lever 3a is operated in a large stroke to increase the opening of the flow control valve 3, the swash plate target position .theta.o is also increased and the modifying coefficient Kr calculated in the block 202 of FIG. 9 takes a larger value (.apprxeq.1), as the tilting amount of the swash |