GB2291986A

GB2291986A - Fluid pressure control system for hydraulic excavators

Info

Publication number: GB2291986A
Application number: GB9515381A
Authority: GB
Inventors: Hee Woo Park
Original assignee: Daewoo Heavy Industries Ltd
Current assignee: Hyundai Doosan Infracore Co Ltd
Priority date: 1994-07-29
Filing date: 1995-07-27
Publication date: 1996-02-07
Anticipated expiration: 2015-07-27
Also published as: GB2291986B; GB9515381D0; JPH0849264A; KR0120281B1

Abstract

A fluid pressure control system for hydraulic excavators includes a variable displacement pump 10 having a swash plate 14 and a regulator unit for changing inclination angle of the swash plate to regulate the fluid discharge volume of the pump. The regulator unit is provided with a) a servo cylinder 36 having a chamber 36c for selective communication with one of the pump and a tank 20 and a chamber 36d for permanent communication with the pump; b) a directional control valve 38 the position of which determines whether the chamber 36c is in communication with the pump oil tank; c) valve 40 kept in abutment to the spool of the valve 38 for shifting the spool into a first position in case of the pump pressure exceeding a predetermined value to thereby reduce the fluid discharge volume of the pump; and d) a load sensing valve 42 mounted in an end-to-end relationship with respect to the valve 40 and responsive to pressure differential between the load pressure and the pump pressure for selectively shifting the spool of the valve 38 to control the fluid discharge volume of the pump. <IMAGE>

Description

FLUID PRESSURE CONTROL SYSTEM FOR HYDRAULIC EXCAVATORS The present invention is generally concerned with a fluid pressure control system for such an industrial vehicle as hydraulic excavator and, more specifically, with a fluid pressure control system for excavators of the type having a simplified discharge volume regulator unit that can adjust the fluid discharge volume of a variable displacement pump in a controlled manner. The invention also lends itself to improve what is referred to as "fine operability" of hydraulic actuators with which the excavators are equipped.

Among hydraulic actuators constituting a typical hydraulic excavator are a swing motor, a boom cylinder, an arm cylinder and a bucket cylinder, all of which can be actuated by virtue of the pressurized working fluid fed from a variable displacement pump via a flow control valve. Driven by a prime mover, e.g., internal combustion engine, the variable displacement pump is adapted to produce a varying amount of the working fluid depending on the magnitude of a load applied to the hydraulic actuators. As used herein, the tern "load" or "load pressure" is intended to mean a resistant force that counteracts the motion vector or moment of the hydraulic actuators caused hy the working fluid.

It is a knowledge of public domain that a load sensing valve is employed to control the fluid discharge volume of the variable displacement pump in a load-dependent fashion. Tlle load sensing valve serves to control the fluid discharge volume in such a manner that the pressure differential between the pump pressure at the exit of the variable displacement pump and the load pressure acting on the hydraulic actuators can be kept at a predetermined value, e.g., 20kg/cm2. Stated in a different way, the fluid discharge volume is so controlled as to increase as the load exerting on the actuators becomes greater and vice versa.

In combination with the load sensing valve, a horsepower control valve or discharge volume restrictor valve is utilized to delimit the maximum fluid discharge volume or permissible horsepower consumption. Mounted separately from the load sensing valve, the horsepower control valve acts to restrict the maximum fluid discharge volume to below a preselected value to thereby prohibit the prime mover from being overloaded. The load sensing valve and the horsepower control valve are both operatively connected to a swash plate of the variable displacement pump through a servo mechanism so as to adjust the inclination angle of the swash plate, which in turn dictates the fluid discharge volume of the variable displacement pump.

Taught in Japanese Pre-examination Patent Publication No. 4-285302 is a fluid pressure control device capable of inhibiting occurence of a discharge volume surge and resultant pressure shock due to the abrupt manipulation of an operating lever or joy stick. The fluid pressure control device includes a variable displacement pump for production of a working fluid with a swash plate attached thereto, a flow control valve adapted to selectively deliver the working fluid to a hydraulic actuator, means responsive to the pressure differential between pump-side pressure and actuator-side pressure for controlling inclination angle of the swash plate to maintain the pressure differential at a predetemlined value, switch means mounted on a manual operating lever for issuing a command signal and means responsive to the command signal for changing the predetermined value.

As another prior art example, Japanese Pre-examination Patent Publication No. 5-150842 discloses a fluid pressure control system for hydraulic excavators that allows the operator to alter the correlation between the stroke of a manual operating lever and the moving speed of a hydraulic actuator without having to detach the operator's hands out of the manual operating lever. A load sensing valve is employed to control a variable displacement pump such that the pressure differential between pump-side fluid pressure and actuator-side pressure can be kept at a predetermined value. The manual operating lever is provided with a push button switch which may be pressed to issue a command signal. In response to the command signal, a controller will cause an electromagnetic valve to become active, thus allowing pilot pressure to act on the load sensing valve.As a result, the predetermined pressure differential varies whereby the correlation between the stroke of the manual operating lever and the moving speed of the hydraulic actuator is subjected to alteration.

Alrhouah the fluid pressure control device and system noted above are effective in suppressing pressure shock which may occur at the time oi abrupt manipulation of the manual operating lever, they have a number of drawbacks as explained below. First, overall construction is complicated and costly due largely to the fact that the load sensing valve, the horsepower control valve and the flow control valve are independently installed at different positions, Secondly, movement of the load sensing valve or the horsepower control valve is directly conveyed to the servo mechanism without going through any damping process, winch may cause an overly clisplacement of the servo mechanism, subjecting the fluid discharge volume to surging and plumbing. Such a fluctuation of the fluid discharge volume will be sustained until and sunless the servo mechanism is made stable. Third, the inability for the prior art device to provide an accurate control of the increase rate of the fluid discharge volume at a minute manipulation interval of the joy stick tends to deteriorate the fine olerability or minute controllability of the hydraulic actuator.

Accordingly, it is an object of the invention to provide a fluid pressure control system for hydraulic excavators that includes a discharbge volume regulator unit of streamlined construction and enhanced reliability and further that can prohibit a swash plate of a variable displacement pump from "overshooting" or unstable back-and-forth movement even when a joy stick is subjected to sudden manipulation, while improving fine opcrability of a hydraulic actuator.

With this object in view, the present invention iwovides a fluid pressure control system for hydraulic excavators including a variable displacement pump for the procluction of a working fluid with puml, pressure, the pump having a swash plate atlilclle(l thercto, at least one hydraulic actuator operable by virtue of the working fluid, a fluid reservoir to which the working fluid is returned for reuse, a now coiitrol valve Ixlsitioned between tlie pump and the actuator for switching off now path of the working fluid, manual operating means for shifting the flow control valve into different positions to control movement of the actuator and a fluid discharge volume regulator unit operatively connected to the pump fi)r changing inclination angle of the swash plate in response to load pressur@ a@ting on the hydrauli@ actuator to regulate fluid discharge volume of the pumlr, characterized in that the fluicl discharge volume regulator unit comprises: a) a servo cylinder ha.ving a servo piston, a piston rod adapted to interconnect the servo piston and the swash plate, a first servo chamber for selective communication with the pump and the fluid reservoir and a second servo chamber for pennanent communication with the pump;; b) a directional control valve having a slx)ol shiftable between a first position in which the first servo chamber comes into communication with the pump and a second position in which the first servo chamber is l)rought into communication with the reservoir;; c) a fluid discharge volume restrictor valve kept in abuttment to the spool of the directional control valve for shifting the spool ilito the first position in case of the pump pressure exceeding a predetemnined valve to allow the working fluid to enter the first servo chamber lo thereby reduce the fluid discharge volume of the pump; and d) a load sensing valve mounted in an end-to-elld relationship with respect to the fluid discharge volume restriclor valve atid responsive to pressure differential between the load pressure and the pump pressure for selectively shifting the spool of the directional control valve into the first and second positions to control the solid discharge volume of the pump.

An example of the present invention will now be described with reference to the accompanying drawings, in which: Figure. 1 is a fluid pressure circuit diagram showing the fluid pressure control system of the invention.

Fig. 2 graphically represents the correlation between the pressure differential acting on a load sensing valve and the fluid discharge volume of a variable displacement pump; Fig. 3 illustrates the interrelation between the pilot fluid pressure exerting on the load sensing valve and tlie input current fed to an electromagnetic proportional control reducing valve; Fig. 4 is a graphical view showing the invot angle of a joy stick versus the pilot fluid pressure applied to the spool of a flow control valve;; Fig. 5 depicts the variation of the aperture area in relation to the spool stroke of line flow control valve; Fig. 6 is a view explaining the corrclation between the pilot fluid pressure applied to the spool of the flow control valve and the fluid quantity delivered to a hydraulic actuator, when the joy stick is subjected to pivotal movement at an initial, fine maiiipulation interval; Fig. 7 shows the relationship between the input current and the pilot fluid pressure for three operation modes of the clectromagnctic proportional control valve chosen by a mode selector switch; and Fig. 8 is a graphical representation illustrating the fading rate of electric current, lhe incrcasing rate or fluid quantity and line rising rate Or pilot fluid pressure over time, as the joy stick is manipulated abruptly to a greater angle.

Referring now to Fig. 1, a fluid pressure control system embodying the invention includes a variable displacement pump 10 which serves to produce a working fluid with pump pressure and an auxiliary pump 12 which is adapted to create a pilot fluid of low pressure, both of the pumps 10, 12 rotatingly driven by means of a prime mover not shown in the drawings for simplicity. The variable displacement pump 10 is provided with a swash plate 14 that can be inclined to various angles to change the quantity of the working fluid discharged by the variable displacement pump 10.

The inclination angle of the swash plate 14 is controlled by a discharge volume regulator unit set forth below such that the larger the inclination angle becomes, the more the working fluid can be discharged from the variable displacement pump 10 and vice versa.

It can be seen in Fig. 1 that the working fluid discharged out of the pump 10 goes to a hydraulic actuator 18 via a main supply line 16. The hydraulic actuator 18 has first and second pressure chambers 18a, 18b, one of which is supplied with the working fluid at a given time, with the other being under drain. Examples of the hydraulic actuator 18 includes a swing motor, a boom cylinder, an arm cylinder and a bucket cylinder. The working fluid drained out of the hydraulic actuator 18 will be gathered in a fluid reservoir 20 and then recycled through the variable displacement pump 10 and the auxiliary pump 12.

A flow control valve 22 is employed to control the movement of the hydraulic actuator 18. As illustrated in Fig. 1, the flow control valve 22 comprises a spool 22a that can be shifted into a first operative position, a second operative position and a neutral position. Use is made of the pilot fluid created by the auxiliary pump 12 to control the position of the spool 22a At the time when the spool 22a is in the first operative position, the working fluid in the main supply line 16 will be admitted into the first pressure chamber 18a of the hydraulic actuator 18, while the working fluid in the second pressure chamber 18b will be drained to the fluid reservoir 20.With the spool 22a placed in the second operative position, the working fluid in the main supply line 16 will be delivered to the second pressure chamber 18b and the working fluid in the first pressure chamber 18a will be retumed back to the fluid reservoir 20. None of the first and second pressure chambers 18a, 18b would be supplied with the working fluid in case of the spool 22a remaining in the neutral position.

First and second pilot pressure chambers 22b, 22c are provided at opposite ends of the flow control valve 22 in such a manner that the first pilot pressure chamber 22b should come into selective communication with the auxiliary pump 12 by way of a first control line 24 and a pilot fluid supply line 26 and that the second pilot pressure chamber 22c should selectively communicate with the auxiliary pump 12 via a second control line 24 and the pilot fluid supply line 26.

Positioned between the pilot fluid supply line 26 and the control lines 24, 28 is a joy stick or manual operating lever 30 that functions to cause the pilot fluid in the pilot fluid supply line 26 to be directed to one of the first pilot pressure chamber 22b, the second pilot pressure chamber 22c and the fluid reservoir 20.

Supplying the pilot fluid into the first pilot pressure chamber 22b by manually operating the joy stick 30 insures that the spool 22a of the now control valve 22 should move to the first operative position i.e., the rightmost position in Fig.

1 to thereby allow the working fluid to enter the first pressure chamber 18a of the hydraulic actuator 18. When the joy stick 30 is in the neutral position, the pilot fluid will be drained to the fluid reservoir 20 without entcring either of the pilot pressure chambers 22b, 22c.

A pressure compensator valve 32 serves to regulate the pressure at which the working fluid is supplied to the hydraulic actuator 18 via the flow control valve 22. The pressure compensator valve 32 is provided with a spool 32a which has the ability to shift among first, second and third operative positions depending on the pressure developed within a load sensing line 34. Should the fluid pressure in the load sensing line 34 overwhelms the load pressure exerting on the hydraulic actuator 18, the spool 32a of the pressure compensator valve 32 will move toward the first operative position, i.e., the leftmost position as shown in Fig. 1, to prohibit the working fluid in the main supply line 16 from entering the hydraulic actuator 18.In the event that the pressure in the load sensing line 34 should become equated with the load pressure acting against the hydraulic actuator 18, the working fluid in the main supply line 16 is allowed to enter the hydraulic actuator 18 due to the movement of the spool 32a of the pressure compensator valve 32 into the second operative position, i.e., the middle position. Altematively, if the fluid pressure in the load sensing line 34 is less than the load pressure exerting against hydraulic actuator 18, the spool 32a of the pressure compensator valve 32 will be caused to move into the third operative position, i.e., the rightmost postion, allowing the load pressure to be delivered to tile load sensing line 34. This means that the fluid pressure in the load sensing line 34 is dctcnnined by the load pressure of the hydraulic actuator 18.

A nuid discharge volume regulator unit is employed to change the inclination angle of the swash plate 14 of the variable displacement pump 10 to thereby control the discharge volume or quantity of the working fluid. As clearly shown in Fig. 1, the fluid discharge volume regulator unit comprises a servo cylinder 36, a directional control valve 38, a discharge volume restrictor valve 40 and a load sensing valve 42, all of which are arranged in an end-to-end relationship with one another along a common axis.

The servo cylinder 36 is so configured to have a servo piston 3Ga, a piston rod 3Gb adapted to operatively interconnect the servo piston 36a and the swash plate 14 of the variable displacement pump 10, a first servo chamber 36c defined at one side of the servo piston 36a for selective communication with either of the variable displacement 10 and the fluid reservoir 20 via the directional control valve 38 and a second servo chamber 36d provided at the other side of the servo piston 36a for permanent communication with the variable displacement pump 10.Introducing the working fluid into the first servo chamber 36c will cause the piston rod 36b to extend such that the inclination angle of the swash plate 14 of the variable displacement pump 10 may become smaller to reduce the discharge volume of the working fluid. To the contrary, if the working fluid is admitted into the second servo chamber 36d to have the piston rod 36b retract, the inclination angle of the swash plate 14 will become greater to increase the discharge volume of the working fluid.

Primary role of the directional control valve 38 is to cause the working fluid to be fed into or exhausted out of the first servo chamber 3Gc of the servo cylinder 36. The directional control valve 38 is provided with a spool 38a which is shiftable between a first position wherein the first servo chamber 36c comes into communication with the variable displacement pump 10 and a second position wherein the first servo chamber 36c is brought into communication with the fluid reservoir 20. Normally, the spool 38a of the directional control valve 38 is kept urged toward the second position by means of a variable constant compression spring 37.

Retained between the piston rod 36b of the servo cylinder 36 and the spool 38a of the directional control valve 38 is a feedback spring 44 whose main function is to, in response to a sudden displacement of the spool 38a of the directional control valve 38, dampen the otherwise unstable movement of the piston rod 36b, helping stabilize the swash plate 14 to thereby shorten the time period taken in reaching a target discharge volume.

The position of the directional control valve 38 continues to be controlled by virtue of the fluid discharge volume restrictor valve 40 and the load sensing valve 42. In the illustrated embodiment, the restrictor valve 40 includes a spool 40a having a pump pressure chamber 40b in communication with the variable displacement pump 10 and a plunger 40c extendibly fitted at one end into the the pump pressure chamber 40b, the other end of the plunger 40c abutting the spool 38a of the directional control valve 38. In case of the fluid pressure in the pump pressure chamber 40b exceeding a predetermined upper limit, the plunger 40c is extended against the spool 38a of the directional control valve 38 to displace the latter into its first position.This results in a decreased inclination angle of the swash plate 14 and a reduced discharge volume of the working fluid, keeping the horsepower consumption in the variable displacement pump 10 free from any increase over a permissible upper limit.

The load sensing valve 42 includes a shuttle piston 42a operatively connected to the spool 40a of the restrictor valve 40, a load pressure chamber 42b in communication with the hydraulic actuator 18 through the load sensing line 34 and a pump pressure chamber 42c communicating with the variable displacement pump 10 via the main supply line 16. In case where the pressure differential between the load pressure in the load pressure chamber 42b and the pump pressure in the pump pressure chamber 42c is less than a preset reference value, thc shuttle piston 42a of the load sensing valve 42 begins to move away from the restrictor valve 40 and the directional control valve 38 whereby the spool 38a of the latter is shifted into the second position as shown in I;ig. 1 to increase the fluid discharge volume of the working fluid. To the contrary, if the pressure differential is greater than the preset reference value the shuttle piston 42a will move toward the restrictor valve 40 and the directional control valve 38 whereby the spool 38a of the latter is shifted into the first position to decrease the fluid discharge volume of the working fluid. This enables the variable displacement pump 10 to produce an optimum amount of the working fluid depending exactly on the severity of the external load applied to the hydraulic actuator 18.

In addition to the above, the load sensing valve 42 may further include a variable pilot pressure chamber 42d that receives the pilot fluid generated by the auxiliary pump 12 via a branch line 46 and an electromagnetic proportional control reducing valve 48. Preferably, the electromagnetic valve 48 is a typical solenoid valve which can open its fluid flow path upon supply of the electric current The aperture area of the electromagnetic valve 48 is adapted to vary with the intensity of the electric current.Stated differently, the aperture area of the electromagnetic valve 48 gets broadened to introduce a greater amount of the pilot fluid into the variable pilot pressure chamber 42d of the load sensing valve 42 as the electromagnetic valve is supplied with a relatively strong electric current At the time of a weak electric current being fed to the electromagnetic valve 48, the aperture area thereof will be narrowed to permit a less amount of the pilot fluid to be admitted into the variable pilot pressure chamber 42d. In a nutshell, the electromagnetic valve 48 can operate as a function of the intensity of the electric current to apply a varying magnitude of thrust force to the shuttle piston 42a of the load sensing valve 42.

To pror,erly control operation of the electromagnetic valve 48, use is made of an electronic control unit which comprises a pressure sensor 50 capable of detecting the pressure of the pilot fluid to be fed to the flow control valve 22 via the first and second control lines 24, 28, a mode selector switch 52 enabling the operator to select a specific operation mode of the electromagnetic valve 48 and a microprocessor 54 responsive to the detected pilot pressure and the selected operation mode for supplying a varying amount of the electric current to the electromagnetic valve 48. It should be appreciated that the pressure sensor 50 is connected to the control lines 24, 28 by way of a shuttle valve 56 for selective fluid communication with one of the control lines 24, 28.

The electromagnetic valve 48 is supplied with an electric current, the intensity of which is in a reverse proportion to the detected pilot pressure.

Even under the circumstance where the pilot pressure is soaring, the electric current will not fade abruptly down to the target value as determined by the detected pilot pressure but will be gradually reduced over time to reach the target value at last By way of using the mode selector switch 52, it becomes possible for the operator to restrict the highest value of the electric current which is supplied to the electromagnetic valve 48. it should be understood that the pressure ol the pilot fluid headed for the pilot pressure chamber 42d through the electromagnetic valve 48 is decreased in proportion to the fading rate of the electric current.As a result, the shuttle piston 42a of the load sensing valve 42 can be caused to move away from the directional control valve 38 at a retarded speed, making it possible to prevent the surging of the working fluid discharge volume even when the joy stick 30 is subjected to a sudden manipulation.

Behavior or operation of the inventive fluid pressure control system will now be described in more detail with reference to Figs. 1 through 7, with emphasis placed on the discharge volume regulator unit.

At the outset, the thrust force of the plunger 40c by which the spool 38a of the directional control valve 38 is urged toward the first position, i.e., in the rightward direction in Fig.1, will be determined by the extension force of the plunger 40c or the pressing force of the shuttle piston 42a acting against the spool 40, whichever is greater. In other words, if the extension force of the plunger 40c is greater than the pressing force of the shuttle piston 42a, the plunger 40c remains apart from the bottom of the pump pressure chamber 40b, which means that the pressing force of the shuttle piston 42a do not affect the movement of the spool 38a of the directional control valve 38.Accordingly, the fluid discharge volume in the variable displacement pump 10 is governed by the restrictor valve 40 rather than the load sensing valve 42. To the contray, if the extension force of the plunger 40c is less that the pressing force of the shuttle piston 42a, the plunger 40c comes into contact with the bottom of the pump pressure chamber 40b such that the spool 38a of the directional control valve 38 can be displaced by the pressing force of the shuttle piston 42a This means that the fluid discharge volume in the variable displacement pump 10 is under control of the load sensing valve 42 alone.

As stated above, the load pressure chamber 42b in the load sensing valve 42 is adapted to receive the load pressure Pl, while the pump pressure chamber 42c thereof is in communication with the variable displacement pump 10 to receive the pump pressure P2. For this reason, the thrust force under which the shuttle piston 40a is biased rightwardly in Fig. 1 will grow in proportion to the pressure differential dP between the load pressure P1 and the pump pressure P2. As the thrust force becomes great enough to overcome the resistant force of the feedback spring 44, the spool 38a of the directional control valve 38 begins to move toward the first position.This insures that the working fluid should be introduced into the first servo chamber 3Gc of the servo cylinder 36 to bias the servo piston 36a and the piston rod 36b leftwardly, making smaller the inclination angle of the swash plate 14 and hence reducing the fluid discharge volume in the variable displacement pump 10.

As the piston rod 3Gb of the servo cylinder 36 continues to move leftwardly in this way, the feedback spring 44 comes to exert a repellent force against the spool 38a of the directional control valve 38, whereby the latter has a tendency to come back to the second position to shut off the introduction of the working fluid into the first servo chamber 36c as ,;llustrated in Fig. 1. The smaller the inclination angle of the swash plate 14 is, the greater the repellent force of the feedback sping 44 becomes. In order to keep the spool 38a of the.

directional control valve 38 in a neutral position, the thrust force pushing the shuttle piston 42a of the load sensing valve 42 rightwardly has to grow in line with the repellent force of the feedback spring 44. This indicates that the less the fluid discharge volume of the variable displacement pump 10 is, the greater the pressure differential dP in the load sensing valve 42 has to become.

Reference is made to Fig. 2 wherein solid line curves show the correlation between the fluid discharge volume Q of the variable displacement pump 10 and the pressure clifferential A P in the load sensing valve 42.

The pilot pressure Pi acting within the pilot pressure chamber 42d of the load sensing valve 42 exhibits a linear increase in proportion to the intensity of the electric current I fed to the electromagnetic valve 48, as best shown in Fig.

3. In other words, supplying the electric current I to the electromagnetic valve 48 will make greater the pilot pressure Pi in the pilot pressure chamber 42d, ensuring that the shuttle piston 42a can be pushed rightwardly with ease even under a decreased pressure differential A P. Referring back to Fig. 2, the broken line curves show the manner in which the nuid discharge volume of the variable displacement pump 10 is reduced in response to the increase of the electric current from iso to Il and Ia Meanwhile, if the spool 38a of the directional control valve 38 is pushed rightwardly due to the increase of the pressure differential dP, the working fluid will be admitted into the first servo chamber 36c of the servo cylinder 36 to urge the piston rod 36b in the left-handed direction. This will results in an increased repellent force of the feedback spring 44, which in turn causes the spool 38a of the directional control valve 38 to be urged leftwardly.Such a returning movement of the spool 38a will shut off the introduction of the working fluid into the first servo chamber 36c in a speedy manner.

Accordingly, the vigorous movement of the piston rod 36c caused by the surging of the pressure differential dP can be stabilized within a shortened period of time, thereby preventing the swash plate 14 of the variable displacement pump 10 from "overshooting" or an undue back-and-forth movement On the other hand, reduction of the pressure differential A P will allow the spool 38a of the directional control valve 38 to move leftwardly such that the working fluid in the first servo chamber 36c can be drained to the fluid reservoir 20, permitting the pistom rod 36b of the servo cylinder to displace in the right-handed direction.In response, the repellent force of the feedback spring 44 becomes smaller to such an extent that the spool 38a of the directional control valve 38 is biased again rightwardly to disconnect the first servo chamber 36c with the fluid reservoir 20. Accordingly, even if the pressure differential A P is subjected to an abrupt decrease, the movement of the piston rod 36b can be stabilized within a short period of time, which assists in suppressing the unstable vibration of the swash plate 14.

As illustrated in Fig. 4, the pilot pressure Pi delivered to one of the control lines 24, 28 through the joy stick 30 is controlled to proportionate the pivoting angle a of the joy stick 30. Shown in Fig. 5 is the aperture area A against the stroke S of the spool 22a of the flow control valve 22 when the pilot pressure Pi is supplied to the pilot pressure chamber 22b or 22c of the flow control valve 22.Since the quantity Q of the working fluid passing through the flow control valve 22 is given by the equation:

the correlation of the pressure differential dP and the fluid quantity Q varies as indicated in the solid line curves in Fig. 2, assuming that the apertrre area A is increased stepwise up to 25%, 50%, 75% and 100or. Fig. 6 represents the interrelation of the pilot pressure Pi and the fluid quantity Q for the respective mode of operation of the electromagnetic proportional control reducing valve 48 at a fine manipulation interval of the joy stick 30. As can be seen in Fig. 6, the control gain of the fluid quantity Q becomes smaller as the electric current I " increases.

It has been the conventional practice that a notch is machined on the spool 22a of the flow control valve 22 to make the control gain of the fluid quantity Q at the fine manipulation interval of the joy stick 30 as small as possible, thus improving the fine operability of the hydraulic actuator 18. To further enhance the fine operability, the instant fluid pressure control system employs the pressure sensor 50 that can detect the pilot pressure Pi in the control line 24 or 28 to apply an electric signal indicative of the pilot pressure Pi to the microprocessor 54. Moreover, the microprocessor 51 is adapted to maximize the electric current " I "fed to the electromagnetic valve 48, when the detected pilot pressure Pi remains low, and to gradually reduce the electric current " I " as the detected pilot pressure Pi becomes high.The electric current " I " will reach the lowest value in case where the detected pilot pressure Pi is determined to be the highest. Reference is made to Fig. 7 wherein the pilot pressure variation is shown in relation to the electric currents Io, Il, and E.

It can be understood in Fig. 6 that the control gain of the fluid quantity is decreased to enhance the fine operability when the pivoting angle of the joy stick 30, i.e., the pilot pressure Pi is relatively small. Furthermore, the fine operability under the third operation mode noted by a solid line Il is superior to that under the first operation mode indicated in dotted lines a, b and a solid line lo. Even if the pivoting angle of the joy stick 30 is made greater abruptly, it will be possible to increase the fluid quantity Q gently and to lessen the shock which would otherwise take place in the hydraulic actuator, by way of gradually decreasing the electric current " I " fed to the electromagnetic valve 48, as illustrated in Fig. 8.

As set forth in detail hereinabove, the instant fluid pressure control system offers a number of advantages over the prior art devices in that the fluid discharge volume regulator unit has a simplified and compact structure with no deterioration of the operational reliability. In addition, the swash plate of the variable displacement pump will not suffer any unstable overshooting movement even when the joy stick is manipulated suddenly at a greater pivoting angle.

Employing the electromagnetic valve and the electronic controller enables the load sensing valve to operate in a controlled manner, which helps improve the fine operability of the hydraulic actuator.

While the invention has been shown and described with reference to a preferred embodiment, it should be apparent to one of ordinary skill that many changes and modifications may be made without departing from the spirit and scope of the invention as defined in the claims.

Claims

CLAIMS:

1. A fluid pressure control system for hydraulic excavators including a variable displacement pump for production of a working fluid with pump pressure, the pump having a swash plate attached thereto, at least one hydraulic actuator operable by virtue or the working fluid, a fluid reservoir to which the working fluid is retumed for reuse, a flow control valve position between the pump and the actuator for switching off now path of the working fluid, manual operating means for shifting the now control valve into different positions to control movement of the actuator and a fluid discharge volume regulator unit operatively connected to the pump for changing inclination angle of the swash plate ill response to load pressure acting on the hydraulic actualor Lo regulate fluid discharge volume of the pump, characterized in that the fluid discharge volume regulator unit comprises:: a) a servo cylinder having a servo piston, a piston rod adapted to interconnect the servo piston and the swash plate, a first servo chamber for selective communication with one of the pump an(l the fluid reservoir and a second servo chamber for pemanent communication with the pump;; b) a directional control valve having a spool shiftable between a first position in which the first selvo chamber comes into communication with the pump alld a second position in which the first servo chamber is brought; into communicalion with the reservoir; c) a fluid discharge volume restrictor valve kept in abutiment to the spool of the directional control valve for shifting the slx)ol into the first position in case of the pump pressure exceeding a predetermined value to allow the working fluid to enter tlie first servo chamber to thereby reduce the fluid discharge volume of the pump; and d) a load sensing valve mounted in an end-to-end relationslii; with respect to the fluid discharge volume restrictor valve and responsive to pressure differential between the load pressure and the pump pressure for selectively shifting the spool of the directional control valve into the first and second positions to control the fluid discharge volume of the pump.

2. A fluid pressure control system as claimed in claim 1, wherein the fluid discharge volume regulator unit futtlier comprises a feedback spn.iig retained between the piston rod of the servo cylinder and the spool Or the (lil-eclional control valve.

3. A fluid pressure control system as claimed in claim 1 or 2, wherein the servo cylinder, the directional control valve, the restrictor valve and the load sensing valve are coaxially arranged with respect to one another.

4. A fluid pressure control system as claimed in any preceding claim, wherein the fluid discharge volume restrictor valve includes a spool having a pump pressure chamber for introduction of the working fluid and a plunger slidably fitted into the pump pressure chamber so as to abut against the spool of the directional control valve at its free end.

5. A fluid pressure control system as claimed in claim 4, wherein the load sensing valve is provided with a shuttle listen abutting against tlie slxx)l of the fluid discharge volume restsictor valve, a load pressure chamber in fluid communication with the hydraulic actuator and a pump pressure chamber in fluid communication with the pump, the shuttle piston adapted, on one hand, to move away from the spool of the directional control valve when the pressure differential between the load pressure and the pump pressure is smaller than a preselected value and, on the other hand, to move toward the spool of the directional control valve when the pressure differential is greater than the preselected value.

6. A fluid pressure control system as claimed in any preceding claim, wherein the fluid discharge volume regulator unit further comprises a feedback spring retained between the piston rod of the servo cylinder and the spool of the directional control valve.

7. A fluid pressure control system as claimed in claim 5, wherein the load sensing valve is further provided with a variable pilot pressure chamber for introducing variable pilot pressure therein to bias the shuttle piSt()fl against the spool of the directional control valve.

8. A fluid pressure control system as claimed in claim 7, fliriher comprising an auxiliary pump for generation of a pilot fluid, a joy stick manually operated Thr feeding the pilot fluid to opposite cnds of the flow control valve to shift the flow control valve into one or the different positions, an electromagnetic proportional control reducing valve for supplying the pilot fluid to the variable pilot pressure chaulll)er of the load sensing valve to urge the shuttle piston against the spool of the directional control valve and an electronic controller for controlling aperture area of the electromagnetic valve as a function of energization time interval and input current intensity.

9. A fluid pressure control system as claimed in claim 8, wherein the electronic controller includes a pressure sensor for detecting pressure of the pilot fluid acting on the now control valve, a mode selector switch for enabling an operation mode of the electromagnetic valve to be selected out of more than one operation modes and a microprncessor responsive to the detected pilot fluid pressure and the selected operation mode for delivering a gradually fading electric current to tile electromagnetic valve for a predeterminecl period of time.

10. A fluid pressure control system for hydraulic excavators substantially as hereinbefore described with reference to the accompanying drawings.