HYDRAULIC VALVE APPARATUS

Abstract: This invention relates to a valve device comprising a flow control valve (8; 8A-8G) including supply passages (11; 11a, 11b) connected to pressure oil supply sources (1, 2), load passages (12; 12a, 12b) connected to an actuator (4) and first meter-in variable throttle portions (14; 14a, 14b) disposed between the supply passages and the load passages for providing an opening in accordance with an operation quantity; first signal passages (18; 16a, 17a, 16b, 17b, 18) disposed downstream of the first variable throttle portion for detecting the load pressure of the actuator; tank passages (13; 13a, 13b) connected to a tank (56); exhaust passages (30; 16b, 17b, 16a, 17a) for connecting the first signal passage to the tank passage; and second variable throttle portions (21; 21a, 21b) disposed in the exhaust passages, for changing the degree of opening in accordance with the operation quantity of the flow regulation valve to generate a control pressure different from the loadd pressure in the first signal passage, wherein the control pressure of the first signal passage is transmitted to the oil pressure supply source through the second signal passage (19). Auxiliary throttle means (22; 22a, 22b) are disposed in the first signal passage (18; 16a, 17a, 16b, 17b, 18),the load pressure detected at the passage portion (15; 15a, 15b) of the first signal passage is reduced and a pressure lower than the load pressure is generated as the control pressure in the first signal passage.

Description: TECHNICAL FIELD The present invention relates to a valve apparatus according to the preamble portion of claim 1. Such an apparatus is shown in DE-A-36 43 110. BACKGROUND ART In a hydraulic drive system for civil engineering and construction machines such as hydraulic excavators and cranes, a flow of a hydraulic fluid supplied from a hydraulic fluid supply source to an actuator is controlled by a valve apparatus including a flow control valve. This type hydraulic drive system uses, as a hydraulic fluid supply source. means for controlling the supply pressure to be held higher a fixed value than the load pressure of the actuator. As disclosed in GB 2195745A, one example of such means is a pump regulator which implements a load sensing system for controlling the pump delivery rate such that the delivery pressure of a hydraulic pump is higher a fixed value than the load pressure. Because the hydraulic fluid is supplied with the load sensing system just at a flow rate required by the actuator, undesired supply of the hydraulic fluid is reduced, which is advantageous in economy, On the other hand. the load sensing system also has the shortcoming that the pump delivery pressure cannot be controlled after the intention of an operator because of its dependency on the load pressure. Therefore, when an inertial load such as a swing of hydraulic excavators is turned, the pump delivery pressure increases up to the setting pressure of a main relief valve irrespective of the amount of a flow control valve operated. This raises the problem that an acceleration of the inertial load is maximized and the operator suffers from a large shock. A known one of valve apparatus for use in the hydraulic drive system implementing the above load sensing system is disclosed in US-A-4 685 295. This disclosed valve apparatus comprises a flow control valve having a supply passage communicating with a hydraulic fluid supply source, a load passage communicating with an actuator, and a first meter-in variable restrictor disposed between the supply passage and the load passage and opened dependent on an operation amount thereof; a first signal passage branched from the load passage downstream of the first variable restrictor and including a restrictor and a check valve allowing a hydraulic fluid to flow toward the load passage; a tank passage communicating with a reservoir tank; a discharge passage for communicating the first signal passage with the tank passage; a second variable restrictor provided in the discharge passage and having its opening variable dependent on the operation amount of the flow control valve to produce in the first signal passage a control pressure different from load pressure; and a second signal passage for leading the control pressure in the first signal passage to the hydraulic fluid supply source, the valve apparatus being featured in further compnsing a third signal passage for connecting the first signal passage to the upstream side of the first variable restrictor at a point between the check valve and the second variable restrictor, and a restrictor disposed in the third signal passage. With that valve apparatus, the pressure upstream of the first variable restrictor is reduced by the restrictor in the third signal passage and then led to the first signal passage. Thus, the reduced pressure is led as the control pressure to the hydraulic fluid supply source to perform the load sensing control, so that the pump delivery pressure may be controlled not depending on the load pressure. Also, by adjusting respective openings of the restrictor in the first signal passage, the restrictor in the second signal passage, and the restrictor in the third signal passage into the appropriate relationship, the dependency on the load pressure can be assured to some extent in a range above the predetermined operation amount, so that the flow rate dependent on the operation amount of the flow control valve is obtained. In the above valve apparatus, however, since the first signal passage is branched from the load passage downstream of the first variable restrictor and includes the restrictor, there occurs a flow of the hydraulic fluid passing from the first signal passage through the restrictor therein to the load passage under a normal condition that the operation amount of the flow control valve is so increased as to secure a predetermined differential pressure across the first variable restrictor. Accordingly, the control pressure which is produced in the first signal passage by reducing the pressure upstream of the first variable restrictor is lower than the pressure upstream of the first variable restrictor, e.g., the pump pressure, but higher than the pressure downstream of the first variable restrictor, i.e., the load pressure. Consequently, the differential pressure between the pressure upstream of the first variable restrictor and the control pressure in the first signal passage becomes smaller than the differential pressure across the first variable restrictor. Thus, if the differential pressure across the first variable restrictor is set to a desired value, the differential pressure between the pressure upstream of the first variable restrictor and the control pressure in the first signal passage would be smaller than the desired value. The hydraulic fluid supply source for the load sensing system receives, as an input signal, the differential pressure between the delivery pressure of the hydraulic pump and the aforesaid control pressure to thereby control the delivery rate of the hydraulic pump such that the above differential pressure becomes equal to a preset target value. Accordingly, the smaller differential pressure between the pressure upstream of the first variable restrictor and the control pressure in the first signal passage implies that the target value must be set to a smaller one. The reduced target value leads to the problem that the control gain is also reduced and hunting is more likely to occur. If the differential pressure across the first variable restrictor is set to a larger value, the aforesaid differential pressure as the input signal to the hydraulic fluid supply source for the load sensing system could be increased. But, the larger differential pressure across the first variable restrictor would increase the pressure loss in the first variable restrictor and would be undesirable from the standpoint of economy. An object of the present invention is to provide a valve apparatus which can control the pump delivery pressure and the drive pressure of an actuator dependent on the operation amount of a flow control valve, and can increase the differential pressure as an input signal to a load sensing system, when the actuator is driven. DISCLOSURE OF THE INVENTION To achieve the above object, the present invention provides a valve apparatus for controlling a flow of a hydraulic fluid supplied from a hydraulic fluid supply source to an actuator, comprising a flow control valve having a supply passage communicating with said hydraulic fluid supply source. a load passage communicating with said actuator, and a first meter-in variable restrictor disposed between said supply passage and said load passage and opened dependent on an operation amount thereof; a first signal passage located downstream of said first variable restrictor and having a passage section for detecting load pressure of said actuator; a tank passage communicating with a reservoir tank; a discharge passage for communicating said first signal passage with said tank passage; and a second variable restrictor provided in said discharge passage and having its opening variable dependent on the operation amount of said flow control valve to produce in said first signal passage a control pressure different from said load pressure, the control pressure in said first signal passage being led to said hydraulic fluid supply source through a second signal passage for controlling the fluid volume supplied by said hydraulic fluid supply source; said valve apparatus comprising auxiliary restrictor means disposed in said first signal passage for reducing the load pressure detected in said passage section of said first signal passage so that a pressure lower than the detected load pressure is produced in said first signal passage as said control pressure. With the present invention thus arranged, since the second variable restrictor having an opening variable dependent on the operation amount of the flow control valve is disposed in the discharge passage. and the auxiliary restrictor means is disposed in the first signal passage, so that the load pressure is adjusted by two restrictors; i.e., the second variable restrictor and the auxiliary restrictor means, to thereby create the control pressure, in the sole operation of the above hydraulic actuator, assuming that the target pressure to be held by the load sensing system implemented with the hydraulic fluid supply source is AP, the opening area of the first variable restrictor is A, the opening area of the auxiliary restrictor means is a1, and the opening area of the second variable restrictor is a2, the port pressure of the load passage, i.e., the drive pressure of the hydraulic actuator, is a function of A, al, a2 and ΔP. Because A and a2 are determined dependent on the operation amount of the flow control valve, the drive pressure can be obtained dependent on the operation amount of the flow control valve. Further, because the hydraulic fluid supply source implements the load sensing system, the pump delivery pressure can also be produced dependent on the operation amount of the flow control valve. In the combined operation of the above hydraulic actuator and other one or more actuators, because a pressure compensating valve for controlling the differential pressure across the first variable restrictor is disposed, the port pressure of the load passage, i.e., the drive pressure of the hydraulic actuator, is a function of A, a1, a2 and ΔP*, assuming that the target pressure to be held by the pressure compensating valve is ΔP*. As with the above case, the drive pressure and the pump delivery pressure can be both obtained dependent on the operation amount of the flow control valve. Accordingly, it is possible to carry out the operation as intended by an operator with higher accuracy for providing superior operability, and to control an acceleration of an inertial load driven by the hydraulic actuator for alleviating the shock perceived by the operator. In addition, with the present invention, since the load pressure is introduced to the first signal passage through the auxiliary restrictor means to create the control pressure, the control pressure is lower than the load pressure, and the differential pressure between the pump delivery pressure and the control pressure is larger than the differential pressure across the first variable restrictor. Therefore, the differential pressure across the first variable restrictor can be set to a normal small value which results in small pressure loss, so that the differential pressure between the pump delivery pressure and the control pressure may be a satisfactorily large value. Consequently, it is possible to increase the control gain of the load sensing system and achieve stable control of the hydraulic pump free from hunting. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a schematic view of a hydraulic drive system incorporating a valve apparatus according to a first embodiment of the present invention. Fig. 2 is a detailed view of a pump regulator used in the hydraulic drive system of Fig. 1. Fig. 3 is a characteristic view showing the relationships between the spool stroke of a flow control valve and the opening areas of a first variable restrictor, a second variable restrictor and a fixed restrictor as developed in the first embodiment. Fig. 4 is a diagram schematically showing a hydraulic system induding a signal passage and a discharge passage established in the first embodiment. Fig. 5 is a vertical sectional view of a valve apparatus according to a second embodiment of the present invention. Fig. 6 is a circuit diagram showing the valve apparatus shown in Fig. 5 in terms of function. Figs. 7 (a) and 7(b) are detailed views of a second variable restrictor and a fixed restrictor provided in the valve apparatus shown in Fig. 5. Fig. 8 is a characteristic view showing the relationships between the spool stroke of a flow control valve and the opening areas of a first variable restrictor, the second variable restrictor and the fixed restrictor as developed in the second embodiment shown in Fig. 5. Fig. 9 is a vertical sectional view of a valve apparatus according to a third embodiment of the present invention. Fig. 10 is a vertical sectional view of a valve apparatus according to a fourth embodiment of the present invention. Fig. 11 is a circuit diagram showing the valve apparatus shown in Fig. 10 in terms of function. Fig. 12 is a vertical sectional view of a valve apparatus according to a fifth embodiment of the present invention. Fig. 13 is a schematic view of a hydraulic drive system incorporating a valve apparatus according to a sixth embodiment of the present invention. Fig. 14 is a vertical sectional view of a valve apparatus according to a seventh embodiment of the present invention. Fig. 15 is a vertical sectional view of a valve apparatus according to an eighth embodiment of the present invention. BEST MODE FOR CARRYING OUT THE INVENTION First Embodiment To begin with, a first embodiment of the present invention will be described below with reference to Figs. 1 through 4. This embodiment pertains to a hydraulic drive system for driving a single-acting actuator. In Fig. 1, the hydraulic drive system of this embodiment comprises a hydraulic fluid supply source made up by a hydraulic pump 1 of variable displacement type and a pump regulator 2 for controlling the displacement volume of the hydraulic pump 1 and constituting a load sensing system, a main relief valve 3 for setting maximum pressure of a hydraulic fluid delivered from the hydraulic pump 1, a single-acting actuator, e.g., a hydraulic motor 4, driven by the hydraulic fluid delivered from the hydraulic pump 1, and a valve apparatus 5 for controlling a flow of the hydraulic fluid supplied from the hydraulic pump 1 to the hydraulic motor 4. The pump regulator 2 controls the displacement volume of the hydraulic pump 1 such that a differential pressure Pd - PLXmax between a delivery pressure Pd of the hydraulic pump 1 and a later-described maximum control pressure PLXmax, or a differential pressure Pd - PLX between the pump (delivery) pressure Pd and a later-described control pressure PLX associated with the hydraulic motor 4 in the case of sole operation of the hydraulic motor 4, is balanced with preset pressure ΔP. In other words, the delivery rate of the hydraulic pump 1 is controlled so as to keep the relationship of Pd = PLXmax + ΔP. The pump regulator 2 is detailed in Fig. 2. The pump regulator 2 comprises an actuator 50 operatively coupled to a swash plate 1a of the hydraulic pump 1 for controlling the displacement volume of the hydraulic pump 1, and a regulating valve 51 operated in response to the differential pressure Pd - PLXmax between the pump pressure Pd and the maximum control pressure PLXmax for controlling operation of the actuator 50. The actuator 50 comprises a double-acting cylinder having a piston 50a having opposite end faces of different pressure receiving areas from each other, and a small-diameter cylinder chamber 50b and a large-diameter cylinder chamber 50c positioned to receive the opposite end faces of the piston 50a, respectively. The small-diameter cylinder chamber 50b is communicated with a delivery line 1b of the hydraulic pump 1 through a line 52, whereas the large-diameter cylinder chamber 50c is selectively communicated with the delivery line 1b through a line 53, the regulating valve 51 and a line 54, or with a reservoir tank 56 through the line 53, the regulating valve 51 and a line 55. The regulating valve 51 has two drive parts 51a, 51b in opposite relation. The pump pressure Pd is loaded to one drive part 51a through a line 57 and the line 54, whereas the maximum control pressure PLXmax is loaded to the other drive part 51b through a signal line 19 as a second signal passage described later. A spring 51c is also disposed in the regulating valve 51 on the same side as the driver part 51b. As the maximum control pressure PLXmax detected by the signal line 19 rises, the regulating valve 51 is shifted leftwardly on the drawing to take an illustrated position. In this state, the large-diameter cylinder chamber 50c of the actuator 50 is communicated with the delivery line 1b, whereupon the piston 50a is moved leftwardly on the drawing because of the difference in pressure receiving area between the opposite end faces of the piston 50a to increase the tilting amount of the swash plate 1a, i.e., the displacement volume of the hydraulic pump 1. As a result, the pump delivery rate is increased to raise the pump pressure Pd. With the pump pressure Pd raised, the regulating valve 51 is returned back rightwardly on the drawing. When the differential pressure Pd - PLXmax reaches a target value determined by the spring 51c, the regulating valve 51 is stopped and the pump delivery rate is kept constant. On the contrary, as the maximum control pressure PLXmax lowers, the regulating valve 51 is shifted rightwardly on the drawing. At this shift position, the large-diameter cylinder chamber 50c of the actuator 50 is communicated with the reservoir tank 56, whereupon the piston 50a is moved rightwardly on the drawing to decrease the tilting amount of the swash plate 1a. As a result, the pump delivery rate is decreased to lower the pump pressure Pd. With the pump pressure Pd lowered, the regulating valve 51 is returned back leftwardly on the drawing. When the differential pressure Pd - PLXmax reaches the target value determined by the spring 51c, the regulating valve 51 is stopped and the pump delivery rate is kept constant. In this manner, the pump delivery rate is controlled such that the differential pressure Pd - PLXmax is held at the target differential pressure determined by the spring 51c. Returning to Fig. 1, the valve apparatus 5 comprises a flow control valve 8 for controlling a flow rate of the hydraulic fluid supplied to the hydraulic motor 4, a pressure compensating valve 9 disposed upstream of the flow control valve 8 for controlling the differential pressure across the flow control valve 8 to supply the hydraulic fluid at a substantially constant flow rate irrespective of fluctuations in the load pressure PL of the hydraulic motor 4 and the pump pressure Pd during the combined operation, a supply passage 11 communicating with the pump 1 through the pressure compensating valve 9, and a load passage 12 capable of communicating with the supply passage 11 and connected to the hydraulic motor 4. The flow control valve 8 comprises a spool made up of a spool section 7a, a spool section 7b and a rod 7c integrally formed together. The spool section 7a has formed therein a first meter-in variable restrictor 14 having an opening variable dependent on the operation amount of the flow control valve 8, i.e., the spool stroke, to disconnect or connect between the supply passage 11 and the load passage 12, and a detection port 15 opened downstream of the first variable restrictor 14 for fluid communication with the load passage 12 to detect the load pressure of the hydraulic motor 4. The valve apparatus 5 also comprises a first signal passage (hereinafter simply referred to as a signal passage) 18 communicating with the detection port 15, a shuttle valve 10 disposed downstream of the signal passage 18, a discharge passage 30 branched from the signal passage 18, and a tank passage 13 communicating with the reservoir tank 56. The spool section 7b of the flow control valve 8 has formed therein a second variable restrictor 21 having an opening variable dependent on the spool stroke to connect or disconnect between the discharge passage 11 and the tank passage 13. The second variable restrictor 21 is configured such that it is opened with a predetermined opening when the flow control valve 8 is in a neutral position, and is closed after opening of the first variable restrictor 14 when the operation amount of the flow control valve 8, i.e., the spool stroke, increases. Further, the signal passage 18 has a fixed restrictor 22 as auxiliary restrictor means disposed between the detection port 15 and the point where the discharge passage 30 is branched from the signal passage 18. The second variable restrictor 21 and the fixed restrictor 22 jointly serve to adjust the load pressure detected by the detection port 15 for creating the control pressure PLX in the signal passage 18. When the second variable restrictor 21 is open, a small amount of the hydraulic fluid flows from the detection port 15 to the tank passage 13 through the signal passage 18 and the discharge passage 30. The load pressure detected by the detection port 15 is reduced by the second variable restrictor 21 and the fixed restrictor 22 so that the control pressure PLX lower than the load pressure PL is produced downstream of the fixed restrictor 22 in the signal passage 18. When the second variable restrictor 21 is closed, there occurs no such a flow of the hydraulic fluid thereby to create the control pressure PLX equal to the load pressure. The shuttle valve 10 serves as higher-pressure selector means for selecting maximum one of control pressures including the control pressure PLX. The selected maximum control pressure PLXmax is passed to a signal line 19 as a second signal passage so that the pump regulator 2 is controlled to regulate the displacement volume of the hydraulic pump 1 for implementation of the load sensing load sensing system, as mentioned above. The valve apparatus 5 further comprises passages 31, 32 for leading inlet pressure Pz of the first variable restrictor 14 and the control pressure PLX to the pressure compensating valve 9, respectively. The pressure compensating valve 9 operates so as to hold differential pressure Pz - PLX between the inlet pressure Pz of the first variable restrictor 14 and the control pressure PLX at substantially constant differential pressure ΔP*. As a result, the differential pressure across the flow control valve 8 is controlled to an almost fixed value. Shift timing of the first and second variable restrictors 14, 21 of the flow control valve 8 and the detection port 15 with respect to the spool stroke, as taken place when the spool of the flow control valve 8 is moved from a neutral position leftwardly in Fig. 1 in the above-described valve apparatus 5, will now be explained with reference to a characteristic graph of Fig. 3 showing the relationship between the spool stroke and the respective opening areas. In Fig. 3, a characteristic line 20a represents the opening area of the second variable restrictor 21, a characteristic line 20b represents the opening area between the detection port 15 and the load passage 12, and a characteristic line 20c represents the opening area of the first meter-in variable restrictor 14. In addition, a characteristic line 20d represents characteristics of the fixed restrictor 22. First, as seen from the characteristic line 20a in Fig.3, when the spool of the flow control valve 8 is in a neutral position, the second variable restrictor 21 is open with a predetermined opening, and the control pressure in the signal passage 18 is equal to the tank pressure. When the spool of the flow control valve 8 is moved leftwardly on the drawing from the above condition, the detection port 15 opens to communicate with the load passage 12 so that the load pressure PL of the hydraulic motor 4 shown in Fig. 1 is led to the detection port 15, as seen from the characteristic line 20b in Fig. 3. In this condition, the second variable restrictor 21 is still open. When the spool of the flow control valve 8 is further moved leftwardly, the first meter-in variable restrictor 14 now opens, whereupon the hydraulic fluid supplied through the pressure compensating valve 9 from the hydraulic pump 1 shown in Fig. 1 is introduced to the hydraulic motor 4 through the supply passage 11, the first variable restrictor 14 and the load passage 12 shown in Fig. 1. As seen from the characteristic line 20a, at the time when the first variable restrictor 14 opens, the second variable restrictor 21 still remains opened, but its opening area has started decreasing. Afterward, the opening area of the first variable restrictor 14 is gradually increased with an increase in the spool stroke, whereas the opening area of the second variable restrictor 21 is gradually decreased. Consequently, downstream of the fixed restrictor 22 in the signal passage 18 shown of Fig. 1, the detected pressure is adjusted by the fixed restrictor 22 and the second variable restrictor 21 to create the control pressure PLX lower than the load pressure PL. The control pressure PLX is passed to the regulating valve 51 (see Fig. 2) of the pump regulator 2 through the shuttle valve 10 and the signal line 19 shown in Fig.3, as mentioned above, whereby the pump 1 is controlled such that the delivery pressure Pd is raised up to a value given by Pd = PLX + ΔP. As a result, the delivery pressure Pd of the hydraulic pump 1 and the port pressure of the load passage 12, i.e., the drive pressure (= load pressure) PL of the hydraulic motor 4 can be controlled as described later. When the spool is further moved from the above condition, the second variable restrictor 21 is dosed as seen from the characteristic line 20a in Fig. 3, and the control pressure PLX equal to the load pressure PL is created in the signal passage 18. This control pressure is passed to the pump regulator 2, whereby the pump 1 is controlled such that the delivery pressure Pd is raised up to a value given by Pd = PLX + ΔP. The hydraulic fluid from the hydraulic pump 1 is supplied to the hydraulic motor4 through the pressure compensating valve 9, the supply passage 11, the first variable restrictor 14 and the load passage 12 for operating the hydraulic motor 4 to drive a working member (not shown). Operation in a range of the spool stroke from opening of the first variable restrictor 14 to closing of the second variable restrictor 21, i.e., in a region S1 in Fig. 3, will be explained below. A hydraulic system including the first variable restrictor 14, the detection port 15, the fixed restrictor 22, the signal passage 18, the discharge passage 30, the second variable restrictor 21 and the tank passage 13 can be schematically depicted as shown in Fig. 4. Supposing now that only the hydraulic motor 4 is driven solely and the pressure compensating valve 9 serving to compensate for the differential pressure ΔP* is not operated and is in a full-open state, the supply pressure, i.e., the pump delivery pressure Pd, is equal to the pressure upstream of the first meter-in variable restrictor 14, i.e., the inlet pressure Pz. Also, owing to the presence of the first variable restrictor 14, the fixed restrictor 22 and the second variable restrictor 21 connected in series to the hydraulic fluid flowing out from the tank passage 13 at a flow rate QT, the relationship among the inlet pressure Pz, the port pressure or the load pressure PL, the control pressure LX and the tank pressure PT is expressed by: Pz > PL > PLX > PT = 0 Let it now be assumed that the opening area of the first variable restrictor 14 is A, the opening area of the fixed restrictor 22 is a1, the opening area of the second variable restrictor 21 is a2, and the hydraulic motor 4 is in a port-blocked state due to the inertial load of a driven member, the flow rate of the hydraulic fluid passing through the first variable restrictor 14 is also QT and, therefore, the following equations hold: QT = C·A Pz - PL QT = C·a1 Pz - PLX QT = C·a2 PLX - PT Pd = Pz = PLX + ΔP Elimination of QT, etc. from the above equations (1) through (4) leads to: PL = {A 2 (a1 2 + a2 2 )<a2 2 (a1 2 + A 2 )}ΔP This can be rewritten to: PL = [{1 + (a2<a1) 2 }/ (a2<a1) 2 (a1 2 /A 2 + 1)]ΔP It will be found from the above equation that the value of the port pressure PL is determined from A, , ΔP, a1 and a2. It will be also found from the equation (4) that the value of the pump delivery pressure Pd is likewise determined from A, ΔP, a1 and a2. When the hydraulic motor 4 and other one or more actuators (not shown) are driven simultaneously, the pressure compensating valve 9 is operated to hold the differential pressure between the pressure Pz upstream of the first variable restrictor 14 and the control pressure PLX at the setting value ΔP*. By replacing Pz - PL in the above equation (1) with Pz - PLX and ΔP in the above equation (4) with ΔP*, therefore, the following equation is obtained: PL = {A 2 (a1 2 + a2 2 )/a2 2 (a1 2 + A 2 )}ΔP* Accordingly, it will be found that during the combined operation, the values of the pump delivery pressure Pd and the port pressure PL are also determined from A, ΔP*, a1 and a2. As will be apparent from the forgoing equations (5) through (7), the drive pressure PL of the hydraulic motor 4, i.e., the port pressure, is a function of the opening areas A and a2 which are determined dependent on the spool stroke of the flow control valve 8. Consequently, in either case of the sole operation of the hydraulic motor 4 or the combined operation of the hydraulic motor 4 and other one or more actuators, there can be obtained the port pressure PL dependent on the operation amount of the flow control valve 8, i.e., the spool stroke. With the first embodiment thus arranged, the flow rate of the hydraulic fluid can be controlled primarily by the opening area A of the first meter-in variable restrictor 14 and, as seen from the equation (6), the maximum value of the port pressure PL can be controlled by the ratio of the opening area a2 of the second variable restrictor 21 to the opening area al of the fixed restrictor 22. Therefore, the pressure control and the flow control both necessary for operation of hydraulic machines can be optimally set by appropriate selection of the opening areas A, a1 and a2. Accordingly, it is possible to carry out the operation as intended by the operator with higher accuracy for providing superior operability, and to control an acceleration of the inertial load driven by the hydraulic motor 4 for alleviating the shock perceived by the operator. Further, in this embodiment, since the load pressure PL is introduced to the signal passage through the fixed restrictor 22 to create the control pressure PLX, there exists the relationship of PL > PLX. In the sole operation of the hydraulic motor 4, the pressure compensating valve 9 is fully opened to give Pd = Pz, and the differential pressure ΔP = Pd - PLX between the pump delivery pressure Pd and the control pressure PLX is larger than the differential pressure ΔP* = Pz - PL across the first variable restrictor 14. It is therefore possible to set the differential pressure across the first variable restrictor 14 to a normal small value which results in the reduced pressure loss, while reserving the differential pressure ΔP at a satisfactorily large value. The regulating valve 51 of the pump regulator 2 receives the differential pressure ΔP between the delivery pressure Pd of the hydraulic pump 1 and the control pressure PLX, as an input signal, to control the delivery rate of the hydraulic pump such that the differential pressure ΔP becomes equal to the fixed value determined by the spring 51c. Accordingly, the smaller differential pressure ΔP implies that the spring 51c must be set to a small setting value. With the setting value reduced, the control gain is so reduced that hunting is more likely to occur. With this embodiment, the differential pressure ΔP as the input signal of the pump regulator 2 can be set to a large value as mentioned above, it is possible to increase the control gain for enabling stable control of the hydraulic pump 1 free from hunting. Moreover, in this embodiment, the control pressure PLX is created from the load pressure PL using two restrictors; the fixed restrictor 22 and the second variable restrictor 21. This results in the advantageous effect that the flow rate of the hydraulic fluid passing through the signal passage 18 and the discharge passage 30 to the reservoir tank 56 can be reduced, and the pressure control can be achieved with smaller energy loss. Although the restrictor 22 is a fixed one in the above first embodiment, it may be a variable one whose opening is variable dependent on the spool stroke of the flow control valve 8 as will be understood from the foregoing equations (5) through (7). This modification can further improve control characteristics. While the spool of the flow control valve 8 comprises the spool sections 7a, 7b and the rod 7c integrally formed together, the rod 7c may be made as a separate member. Alternatively, the spool sections 7a, 7b may be arranged to be movable independently and driven by a common pilot pressure. In addition, either one or both of the first and second variable restrictors 14, 21 may be in the form of a poppet valve. Second Embodiment A second embodiment of the present invention will be described with reference to Figs. 5 through 8. This embodiments provides a valve apparatus for driving a double-acting actuator. Fig. 5 is a vertical sectional view of the valve apparatus, and Fig. 6 is a circuit diagram showing the valve apparatus in terms of function. In these drawings, the identical components to those shown in Fig. 1 are denoted by the same reference numerals. In Figs. 5 and 6, a valve apparatus 5A of this embodiment comprises a block 6 forming a body, a flow control valve 8A having a spool 7 slidable in a spool bore 6a defined in the block 6, a pressure compensating valve 9 provided upstream of the flow control valve 8A to control differential pressure between inlet pressure Pz and outlet pressure PL of the flow control valve 8A, i.e., differential pressure Pz - PL across the flow control valve 8A, and a shuttle valve 10 provided downstream of the flow control valve 8A. The block 6 has formed therein two supply passages 11a, 11b communicating with a hydraulic pump 1, two load passages 12a, 12b capable of communicating with the supply passages 11a, 11b, respectively, and connected to a hydraulic actuator shown in Fig. 6, e.g., a swing motor 4A for driving a swing of a hydraulic excavator, and two tank passages 13a, 13b capable of communicating with the load passages 12a, 12b, respectively. The spool 7 has two first meter-in variable restrictors 14a, 14b for communicating the supply passage 11a with the load passage 12a and communicating the supply passage 11b with the load passage 12b, respectively, and being opened dependent on the stroke of the spool 7, two detection ports 15a, 15b capable of being open to the load passages 12a, 12b downstream of the first variable restrictors 14a, 14b, respectively, to detect the load pressure PL of the swing motor 4A, two passages 16a, 16b communicating with the detection ports 15a, 15b, respectively, and two passages 17a, 17b communicating with the passages 16a, 16b, respectively. The block 6 further has a passage 18 capable of communicating with the passages 17a, 17b. The spool 7 is also formed with a second variable restrictor 21a positioned between the passage 17b and the passage 18 and having its opening area variable dependent on the stroke of the spool 7 when the spool 7 is moved rightwardly on the drawing, a second variable restrictor 21b positioned between the passage 17a and the passage 18 and having its opening area variable dependent on the stroke of the spool 7 when the spool 7 is moved leftwardly on the drawing, a fixed restrictor 22a positioned between the passage 17a and the passage 18 and carrying out its function when the spool 7 is moved rightwardly on the drawing, and a fixed restrictor 22b positioned between the passage 17b and the passage 18 and carrying out its function when the spool 7 is moved leftwardly on the drawing. As with the first embodiment, the second variable restrictors 21a, 21 bare configured such that they are open at a predetermined opening when the spool 7 is in a neutral position, and are dosed after opening of the first variable restrictors 14a, 14b when the spool stroke is increased. The detection port 15a, the passages 16a, 17a and the passage 18 jointly constitute a first signal passage for detecting the load pressure of the swing motor 4A downstream of the first variable restrictor 14a, when the spool 7 is moved rightwardly on the drawing. The detection port 15b, the passages 16b, 17b and the passage 18 jointly constitute a first signal passage for detecting the load pressure of the swing motor 4A downstream of the first variable restrictor 14b, when the spool 7 is moved leftwardly on the drawing. Further, the detection port 15b and the passages 17b, 16b jointly constitute a discharge passage for communicating the first signal passage 15a, 16a, 17a, 18 established when the spool 7 is moved rightwardly on the drawing, with the tank passage 13b, the second variable restrictor 21a being disposed in this discharge passage. The detection port 15a and the passages 17a, 16a jointly constitute a discharge passage for communicating the first signal passage 15b, 16b, 17b, 18 established when the spool 7 is moved leftwardly on the drawing, with the tank passage 13a, the second variable restrictor 21b being disposed in this discharge passage. The fixed restrictor 22a is disposed in the first signal passage 15a, 16a, 17a, 18 established when the spool 7 is moved rightwardly on the drawing, and serves as auxiliary restrictor means for reducing the load pressure detected by that first signal passage to create the control pressure PLX lower than the load pressure. The fixed restrictor 22b is disposed in the first signal passage 15b, 16b, 17b, 18 established when the spool 7 is moved leftwardly on the drawing, and serves as auxiliary restrictor means for reducing the load pressure detected by that first signal passage to create the control pressure PLX lower than the load pressure. The control pressure PLX produced in the passage 18 constituting a part of the first signal passage is, similarly to the first embodiment, introduced to a signal line 19 as a second signal passage through the shuttle valve 10 as higher-pressure selector means, and used for the load sensing control by the pump regulator 2. The second variable restrictors 21a, 21b and the fixed restrictors 22a, 22b are detailed in Figs. 7(a) and 7(b). Of these drawings, Fig. 7(a) shows a neutral state of the spool 7, and Fig. 7(b) shows a state in which the spool 7 has been moved leftwardly. Arrows in Fig. 7(b) indicate a flow of the hydraulic fluid in the signal passage and the discharge passage. Shift timing of the first and second variable restrictors 14a, 14b and 21a, 21b and the detection ports 15a, 15b with respect to the spool stroke of the flow control valve 8A is shown in Fig. 8. Characteristics of the first variable restrictors 14a, 14b, i.e., the relations of their opening areas with respect to the stroke of the spool 7, are set identical to the characteristic line 20c in Fig. 3. Characteristics of the second variable restrictors 21a, 21b are set identical to the characteristic line 20a in Fig. 3. Characteristics of the fixed restrictors 22a, 22b are set identical to the characteristic line 20d in Fig. 3. The opening areas between the detection ports 15a, 15b and the load passages 12a, 12b are set identical to the characteristic line 20b in Fig. 3. In addition, the characteristic line 20e indicates the opening area between the detection ports 15a, 15b and the tank passages 13a, 13b. The swing motor 4A is a double-acting actuator. In a main line connected to the load passages 12a, 12b of the valve apparatus 5A, there is disposed a counter balance valve 35 for blocking off the holding pressure produced when the swing (not shown) is installed on a slope. With the second embodiment thus arranged, when the spool 7 is moved from a neutral position rightwardly in Fig. 5 with an intention of driving the swing motor 4A solely, the communication between the detection port 15a and the tank passage 13a is first cut off as seen from the characteristic line 20e in Fig. 8. When the spool 7 is further moved from the above condition, the first variable restrictors 14a, 14b, the second variable restrictors 21a, 21b, the fixed restrictors 22a, 22b, the detection ports 15a, 15b, and the load passages 12a, 12b, though each provided in pair, exhibit their characteristics identical to those in the foregoing first embodiment. Accordingly, the above-described equations (5) through (7) are also satisfied in the second embodiment. As a result, the port pressure, i.e., the drive pressure PL, and the delivery pressure Pd of the hydraulic pump 1 can be controlled dependent on the operation amount of the flow control valve 8A, i.e., the spool stroke, thereby providing the similar advantageous effect to that in the first embodiment. Because the control pressure PLX created in the passage 18 through the fixed restrictors 22a, 22b meets the relationship of PL > PLX, the differential pressure ΔP = Pd - PLX between the pump delivery pressure Pd and the control pressure PLX can be a satisfactorily large value. Further, because the control pressure PLX is created using two restrictors; the fixed restrictor 22a, 22b and the second variable restrictor 21a, 21b, the flow rate of the hydraulic fluid passing from the detection port 15a, 15b as the signal passage to the tank passage 13b, 13a through the passage 18 and the detection port 15b, 15a as the discharge passage can be reduced, and the pressure control can be achieved with smaller energy loss. In this point, the similar advantageous effect to that in the first embodiment can also be obtained. It is of course a matter that although the restrictors 22a, 22b are fixed ones in this embodiment, they may be variable ones whose openings are variable dependent on the stroke of the spool 7, as with the foregoing first embodiment. Third Embodiment A third embodiment of the present invention will be described with reference to Fig. 9. This embodiment is to give the valve apparatus with a function of reserving the holding pressure of the actuator. In Fig. 9, a valve apparatus 5B of this embodiment has second variable restrictors 21a, 21b and fixed restrictors 22a, 22b identical to those in the foregoing second embodiment. A check valve 23 with small spring pressure is slidably fitted in a spool 7 which constitutes a flow control valve 8B. When the spool 7 is in the vicinity of a neutral position, the passage 16a is connected to the tank passage 13a through the check valve 23, thereby forming the discharge passage. When the spool 7 is moved rightwardly on the drawing, the fixed restrictor 22a functions between the detection port 15a and the passage 18, and the supply passage 11a is communicated with the load passage 12a through the check valve 23 upon opening of the first meter-in variable restrictor 14a. When the spool 7 is moved leftwardly on the drawing, the passage 18 is communicated with the tank passage 13a through the second variable restrictor 21b, the passage 17a, the passage 16a and the check valve 23 which jointly define the discharge passage. Then, as an actuator of which operation is controlled by the valve apparatus 5B, there is provided a hydraulic cylinder, e.g., a boom cylinder 4B for driving a boom of hydraulic excavators. The boom cylinder 4B is communicated at the head side with the load passage 12a in which the check valve 23 is located, and at the rod side with the load passage 12b. During operation of a boom (not shown) carried out by the boom cylinder 4B, for example, when the boom is held at an elevated level in air, the dead load of the boom acts on the boom cylinder 4B and the holding pressure is produced in the head side line of the boom cylinder 4B, i.e., the load passage 12a. With the third embodiment thus arranged, when the spool 7 of the flow control valve 8B is moved rightwardly with an intention of driving the boom cylinder 4B solely, the detection port 15a is first disconnected from the tank passage 13a, and the detection port 15a is then communicated with the load passage 12a. Afterward, the passage 16a is communicated with the supply passage 11a through the first meter-in variable restrictor 14a. Consequently, the first variable restrictor 14a, the fixed restrictor 22a and the second variable restrictor 21a now constitute the foregoing hydraulic system shown in Fig. 4. As a result, the above-described equations (5) through (7) are also satisfied in the third embodiment, whereby the port pressure PL and the pump delivery pressure can be controlled dependent on the spool stroke of the flow control valve 8B as with the foregoing second embodiment. At this time, the hydraulic fluid is supplied from the supply passage 11a to the head side of the boom cylinder 4B through the first variable restrictor 14a, the passage 16a, the check valve 23 and the load passage 12a. In this connection, if the aforesaid holding pressure is produced in the head side line of the boom cylinder 48, i.e., the load passage 12a, the pressure in the passage 16a is determined by the stroke of the spool within the stroke range where the hydraulic system shown in Fig. 4 is established, and that pressure may be lower than the holding pressure produced in the load passage 12a. To cope with that, in this embodiment, the check valve 23 acts to prevent the hydraulic fluid from flowing from the load passage 12a to the passage 16a. Therefore, even if the holding pressure is produced in the head side line of the boom cylinder 4B, i.e., the load passage 12a, the hydraulic fluid under pressure held in the load passage 12a will not flow into the passage 16a and then flow out to the reservoir tank through the fixed restrictor 22a, the passage 18, the second variable restrictor 21a, and the discharge passage which is defined by the passages 17b, 16b and the detection port 15b. Consequently, this embodiment can reserve a holding function to prevent contraction of the boom cylinder 4B, i.e., a drop of the boom by the gravity or dead load. On the contrary, when the spool 7 of the flow control valve 8 is moved leftwardly, the supply passage 11b is communicated with the load passage 12b in which no holding pressure occurs, through the first meter-in variable restrictor 14b and the passage 16b. Also, the second variable restrictor 21a, the passages 17a, 16a, the check valve 23 and the detection port 15a jointly define the discharge passage led to the tank passage 13a. In this embodiment, therefore, since the hydraulic system shown in Fig. 4 is established by the fixed restrictor 22b and the second variable restrictor 21b, the foregoing equations (5) through (7) are satisfied and the port pressure PL and the pump delivery pressure can be controlled desirably. At this time, the returning hydraulic fluid on the head side of the boom cylinder 4B is discharged from the load passage 12a to the tank passage 13a through the passages 24, 16a and the check valve 23. Thus, with satisfaction of the foregoing equations (5) through (7), the third embodiment can control the port pressure (drive pressure) PL and the pump delivery pressure dependent on the spool stroke of the flow control valve 8B, and can achieve force control for regulating thrust of the boom cylinder 4B with the control of the port pressure. In addition, since the third embodiment indudes the check valve 23 between the load passage 12a and the first variable restrictor 14a, when the spool 7 shown in Fig. 9 is moved rightwardly to extend the boom cylinder 4B, the hydraulic fluid held under pressure on the head side of the boom cylinder 4B will not flow into the passage 16a, and the boom (not shown) can be prevented from dropping by the dead load upon contraction of the boom cylinder 4B. Fourth Embodiment Afourth embodiment of the present invention will be described with reference to Figs. 10 and 11. This embodiment is to provide a valve apparatus for use in a double-acting actuator which has no counter balance valve. In Fig. 10, a valve apparatus 5C indudes a pair of check valves 25a, 25b disposed in a spool 7 of the flow control valve 8C. The check valve 25a is disposed between the supply passage 11a and the load passage 12a as well as the tank passage 13a, while the check valve 25b is disposed between the supply passage 11b and the load passage 12b as well as the tank passage 13b. A swing motor 4A having no counter balance valve is provided as an actuator to drive a swing (not shown). The spool 7 of the flow control valve 8C is depicted as shown in Fig. 11 in terms of function. When the spool 7 is moved rightwardly from the condition shown in Fig. 11, a region S1 of this spool 7 corresponds to the aforesaid region S1 in Fig. 8, i.e., the stroke region where the fixed restrictor 22a and the second variable restrictor 21a both function as restrictors. Also, a region S2 of the spool 7 shown in Fig. 11 corresponds to the aforesaid region S2 in Fig. 8, i.e., the stroke region where the second variable restrictor 21a is closed. The remaining structure of the valve apparatus 5C is identical to that shown in Fig. 9. With the fourth embodiment thus arranged, when the spool 7 of the flow control valve 8C is moved rightwardly in Figs. 10 and 11, for example, the hydraulic system shown in Fig. 4, which indudes the first variable restrictor 14a, the fixed restrictor 22a, and the discharge passage having the second variable restrictor 21a and the check valve 25 therein, is established in a range of the region S1 shown in Fig. 11. Therefore, the foregoing equations (5) through (7) are satisfied and the port pressure PL can be controlled dependent on the stroke of the spool 7, i.e., the lever operation amount of the flow control valve 8C in any operation of driving the swing motor solely or in combination with other one or more actuators. This is equally applied to the case where the spool 7 is moved leftwardly in Figs. 10 and 11. As a result, the similar advantageous effect to that in the foregoing second embodiment can be obtained. Furthermore, if the swing (not shown) is installed on a slope, for example, the holding pressure is produced in either the load passage 12a or 12b both connected to the swing motor 4A. In such a case, when the spool 7 of the flow control valve 8C is moved, the hydraulic system shown in Fig. 4 is established in a range of the region S1 shown in Fig. 11 as mentioned above, and the pressure in the passage 16a or 16b is determined by the stroke of the spool 7, resulting in that the pressure in the passage 16a or 16b may be lower than the holding pressure produced in the load passage 12a, 12b. With this fourth embodiment, however, no matter which one of the load passages 12a, 12b is subjected to the holding pressure, the hydraulic fluid held in the load passage under pressure is prevented from flowing into the supply passage 11a, 11b by the corresponding one of the check valves 25a, 25b. This ensures it to avoid operation of the swing motor 4A not intended by the operator, i.e., undesired motion of the swing (not shown). Fifth Embodiment A fifth embodiment of the present invention will be described with reference to Fig. 12. This embodiment has an operator check, in place of the check valve, to block off the holding pressure. In Fig. 12, a valve apparatus 5D of this embodiment has an operator check 26 in a load passage 12a which is defined in a block 6 constituting the valve apparatus body and is subjected to the holding pressure of a boom cylinder 4B. The remaining structure is identical to that of the third embodiment shown in Fig. 9. With this fifth embodiment thus arranged, the foregoing equations (5) through (7) are satisfied on the basis of the hydraulic system induding the first variable restrictors 14a, 14b, as well as the corresponding fixed restrictors 22a, 22b and the second variable restrictors 21a, 21b. Therefore, the port pressure PL and the pump delivery pressure can be controlled dependent on the lever operation amount of the flow control valve 8B. In addition, when the hydraulic fluid is supplied to the load passage 12a to extend the boom cylinder 4B, the operator check 26 is opened only after the pressure in the load passage 12a becomes larger than the holding pressure acting on the head side of the boom cylinder 4B, allowing the hydraulic fluid to be supplied to the head side of the boom cylinder 4B for driving of the boom cylinder 4B. Consequently, the hydraulic fluid boosted in pressure for holding the boom cylinder 4B is prevented from flowing into the supply passage 11a, and the similar advantageous effect to that in the third embodiment of Fig. 9 can be obtained. Sixth Embodiment A sixth embodiment of the present invention will be described with reference to Fig. 13. A valve apparatus 5E according to the sixth embodiment, shown in Fig. 13, has a limiter 36 for limiting the operation amount of a flow control valve 8E to a predetermined amount in short of the maximum stroke, in addition to the structure of the foregoing first embodiment shown in Fig. 1. The limiter 36 comprises, for example, a projection against which a spool section 7a of the flow control valve 8E strikes for restriction of its movement A maximum value of the stroke restricted by the limiter 36 corresponds to a point X contained in the region S1 of Fig. 3 by way of example. The sixth embodiment thus arranged is effective in the case where the inertial load to be driven by the hydraulic motor 4 is relatively small and, therefore, the load pressure is small. The installed position of the limiter 36 is previously set such that when the flow control valve 8E is operated until the spool section 7a strikes against the limiter 36, the load pressure PL determined by the foregoing equations (5) through (7) has a value substantially in agreement with the drive pressure necessary for the hydraulic motor 4. With such presetting, the maximum port pressure is determined from the above equation (6), and the load pressure applied to the hydraulic motor is limited to the relatively small load pressure PL corresponding to the point X in Fig. 3. Accordingly, with this sixth embodiment, since the basic structure is identical to that of the foregoing first embodiment, the aforesaid equations (5) through (7) are satisfied, whereby the flow rate and the load pressure PL can be controlled as intended by the operator. In addition, without the need of especially installing a relief valve adapted to release the surplus load pressure produced in a circuit containing the hydraulic motor 4, it is possible to protect equipment in that circuit from damage, and to suppress energy loss which would otherwise be caused with release of the surplus load pressure, resulting in an advantage of economy. Seventh Embodiment A seventh embodiment of the present invention will be described with reference to Fig. 14. A valve apparatus 5F according to the seventh embodiment, shown in Fig. 14, has a limiter 36A in addition to the structure of the foregoing second embodiment shown in Fig. 5. The limiter 36Acomprises a screw 37 for limiting the stroke of a spool 7 of a flow control valve 8F to a predetermined position in short of the maximum stroke, and a lock nut 38 for fastening the screw 37 in place. As with the foregoing sixth embodiment, this seventh embodiment can also limit the drive pressure of the actuator to be controlled by the valve apparatus 5F, and provide the similar advantageous effect to that in the sixth embodiment. Eighth Embodiment An eighth embodiment of the present invention will be described with reference to Fig. 15. A valve apparatus 5G according to the eighth embodiment has a pilot valve 39 and a pressure reducing valve 36B for reducing pilot pressure generated by the pilot valve 39. The pressure reducing valve 36B serves as a limiter for limiting the operation amount of a spool 7 of a flow control valve 8G. The remaining structure is identical to that of the foregoing second embodiment shown in Fig. 5. Thus, by adjusting the pilot pressure, it is also possible to achieve the similar operation to that in the foregoing seventh embodiment, and to provide the similar advantageous effect to that in the seventh embodiment. With the pressure reducing valve 36B as a limiter being in the form of a solenoid proportional valve, the maximum pilot pressure, i.e., the maximum stroke, can be adjusted using an electric signal. INDUSTRIAL APPLICABILITY According to the present invention, when the flow control valve is operated from a neutral position in the sole or combined operation of one or more actuators, the delivery pressure of the hydraulic pump and the drive pressure of the actuator can be controlled dependent on the operation amount of the flow control valve. This reliably eliminates the event that the pump delivery pressure may be increased up to the setting pressure of a main relief valve against the intention of an operator, and ensures excellent operability. Also, the control of the drive pressure permits force control of the actuator so that, when the actuator drives an inertial load, an acceleration of the inertial load may be controlled. As a result, the shock perceived by the operator can be alleviated. Further, since the load pressure is reduced by a fixed restrictor to create the control pressure, the differential pressure between the pump delivery pressure and the control pressure can be set to a satisfactorily large value to thereby enable the loading sensing control of the hydraulic pump free from hunting. In addition, since the control pressure is created using two restrictors; i.e., the fixed restrictor and the second variable restrictor, the flow rate of the hydraulic fluid flowing from the signal passage to the reservoir tank through the discharge passage can be reduced so as to achieve the pressure control with small energy loss.

Claims: A valve apparatus (5A-5G) for controlling a flow of a hydraulic fluid supplied from a hydraulic fluid supply source (1, 2) to a double-acting actuator (4A; 4B), comprising a flow control valve (8A-8G) having supply passages (lla, llb) communicating with said hydraulic fluid supply source, a pair of load passages (12a, 12b) communicating with said actuator, and a pair of first meter-in variable restrictors (14a, 14b) disposed between said supply passages and said pair of load passages, respectively, and opened alternatively dependent on an operating direction of the flow control valve (8A-8G) to an opening dependent on an operation amount of the same; a pair of first signal passages (16a, 17a, 16b, 17b, 18) located downstream of said pair of first variable restrictors, respectively, and having passage sections (15a, 15b) for detecting load pressure of said actuator alternatively dependent on the operating direction of said flow control valve (8A-8G) ; a pair of tank passages (13a, 13b) each communicating with a reservoir tank (56); a pair of discharge passages (16b, 17b, 16a, 17a) for communicating said pair of first signal passages with said pair of tank passages, respectively; a pair of second variable restrictors (21a, 21b) provided in said pair of discharge passages, respectively, and having their openings variable dependent on the operation amount of said flow control valve to produce in said pair of first signal passages a control pressure different from the load pressure detected in the corresponding first signal passage, alternatively dependent on the operating direction of said flow control valve (8A-8G) and a pair of auxiliary restrictor means (22a, 22b) disposed in said pair of first signal passages (16a, 17a, 16b, 17b, 18), respectively, for reducing the load pressure detected alternatively in said passage sections (15a, 15b) of said pair of first signal passages so that a pressure lower than the detected load pressure is produced in the corresponding first signal passage as said control pressure, the control pressure produced alternatively in said pair of first signal passages being led to said hydraulic fluid supply source through a second signal passage (19) for controlling the fluid volume supplied by said hydraulic fluid supply source; wherein said valve apparatus includes a valve block (6), and a spool (7) inserted in said valve block (6) movably in its axial direction to provide said flow control valve (8A-8G), said pair of first variable restrictors (14a, 14b) being formed by said spool (7), said spool (7) having a pair of inner passages (16a, 16b); characterized in that said valve apparatus further comprises: a single pressure compensating valve (9) disposed in said valve block (6) outside said spool (7) at a location upstream of said pair of first variable restrictors (14a, 14b) for controlling a differential pressure across said pair of first variable restrictors (14a, 14b); a single passage (18) formed in said valve block (6) around said spool (7) at an axially central portion thereof and communicating with said pressure compensating valve (9) to provide said control pressure thereto; a pair of variable restrictors (21a, 21b) and a pair of fixed restrictors (22a, 22b) formed by said outer periphery of said spool (7) on opposed sides of said single passage (18), respectively, so as to communicate therewith; said pair of inner passages (16a, 16b) formed in said spool (7) include a pair of passages (17a, 17b) opening said pair of fixed restrictors (22a, 22b), respectively; one (16a, 17a) of said pair of inner passages (16a, 16b, 17a, 17b) and said single passage (18) functioning as one of said pair of first signal passages and the other (16b, 17b) of said pair of inner passages functioning as one of said pair of discharge passages whilst one (22a) of said pair of fixed restrictors (22a, 22b) positioned on the side of said one inner passage (16a, 17a) functioning as one of said pair of auxiliary restrictors and one (21a) of said pair of variable restrictors (21a, 21b) positioned on the side of said other inner passages (16b, 17b) functioning as one of said pair of second variable restrictors when one (14a) of said pair of first variable restrictors (14a, 14b) is opened upon said spool (7) axially moving in one direction, one (16a, 17a) of said pair inner passages (16a, 16b, 17a, 17b) functioning as the other of said pair of discharge passages and the other (16b, 17b) of said pair of inner passages and said single passage (18) functioning as the other of said pair of first signal passages whilst the other (22b) of said pair of fixed restrictors (22a, 22b) positioned on the side of said other inner passage (16b, 17b) functioning as the other of said pair of auxiliary restrictors and the other (21b) of said pair of variable restrictors (21a, 21b) positioned on the side of said one inner passages (16a, 17a) functioning as the other of said pair of second variable restrictors when the other (14b) of said pair of first variable restrictors (14a, 14b) is opened upon said spool axially moving in the other direction. A valve apparatus according to claim 1, wherein said pair of second variable restrictors (21; 21a, 21b) are configured to be open to a predetermined opening when said flow control valve (8; 8A-8G) is in a neutral position, and closed after opening of said pair of first variable restrictors when said flow control valve is operated. A valve apparatus according to claim 1, further comprising a shuttle valve (10) disposed in said single passage (18) for selecting maximum one of control pressures including the control pressure produced in said pair of first signal passages (18; 16a, 17a, 16b, 17b, 18), and leading the selected maximum pressure as the control pressure to said second signal passage (19). A valve apparatus according to claim 1, wherein an operator check (26) is disposed in said pair of load passages (12a, 12b). A valve apparatus according to claim 1, further comprising limiter means (36; 36A, 36B) for limiting the operation amounf of said flow control valve (8E; 8F; 8G) to a predetermined value. A valve apparatus according to claim 1, wherein said pair of inner passages have first passage sections (16a, 16b) positioned downstream of said pair of first variable restrictors (14a, 14b) and second passage sections (15a, 15b) capable of communicating said pair of load passages (12a, 12b) with said pair of tank passages (13a, 13b), respectively, and check valves (25a, 25b) are disposed between said first passage sections and said second passage sections, respectively, for allowing the hydraulic fluid to flow only in a direction toward said second passage sections from first passage sections.

Inventors: HIRATA TOICHI, KAJITA YUSUKE, SUGIYAMA GENROKU, HIRATA, TOICHI, KAJITA, YUSUKE, SUGIYAMA, GENROKU
Applicants: HITACHI CONSTRUCTION MACHINERY, HITACHI CONSTRUCTION MACHINERY CO., LTD.
Application year: 1990
Main classification: F15B11
IPC codes: E02F9, F15B13