JPH07109205B2 - Hydraulic control valve - Google Patents

Description

TECHNICAL FIELD The present invention relates to a hydraulic control valve suitable for a system in which a plurality of hydraulic actuators are driven by a single pump.

[Conventional technology and its technical problem]

In a construction machine, a single large-capacity hydraulic pump is used, and a plurality of actuators are supplied with oil discharged from these pumps.For example, in a para excavator, a turning hydraulic motor, a left traveling motor, a right traveling motor, a boom cylinder, It is common to drive arm cylinders and bucket cylinders.

As a control means of this system, a plurality of directional switching valves are connected between each actuator and the hydraulic pump, and the amount of oil flowing through the directional switching valve is pressure-compensated so that the operating speed of the actuator does not change due to load fluctuations. Compensation is being done with a valve. As this pressure compensation method,
In 60-11706, the throttle opening of the pressure compensation valve is
It has been proposed that the pressure is controlled by the pressure difference between the pump discharge pressure and the signal pressure from the shuttle valve, instead of the spring force of a general-purpose pressure compensation valve. That is, in this prior art, a pressure compensating valve is provided upstream of the switching valve, the closing side of the pressure compensating valve is loaded with the pressure reaching the switching valve, and the opening side is loaded with the load pressure of the actuator. On the other hand, the closing side of the pressure compensation valve is loaded with the maximum load pressure of the actuator in operation selected by the shuttle valve, and the opening side is loaded with the discharge pressure of the pump. Then, this prior art (hereinafter referred to as prior art 1) is embodied, and a switching valve, a pressure compensating valve and a shuttle valve are skillfully incorporated into one valve body, and the same degree as a normal multiple control valve without a pressure compensating valve. The control valve with the size of
It is proposed in Japanese Patent Publication No. 150201 (hereinafter referred to as Prior Art 2).

However, in the prior art 2, although the pressure compensating valve and the shuttle valve are incorporated in the valve body of the switching valve, there is still a problem in that the structure of the load pressure introducing passage to the pressure compensating valve and the shuttle valve is complicated. It was Further, in both the prior art 1 and the prior art 2, the load pressure of the actuator and the pressure upstream of the cutout of the switching valve are opposed to each other, while the pump discharge pressure and the maximum load pressure selected by the shuttle valve are opposed to each other. The throttle opening of the pressure compensation valve was controlled by the pressure difference. For this reason, the degree of freedom in controlling the throttle opening of the pressure compensating valve is poor, and it has been difficult to respond individually to various demands of operating conditions of a plurality of actuators.

As a countermeasure against this, an improved control valve is proposed in Japanese Patent Laid-Open No. 2-134402 (hereinafter referred to as the third prior art).
However, also in this third prior art, the balance piston of the pressure compensating valve has the opening-side first pressure receiving surface facing the load pressure and the opening-side second pressure receiving surface contacting the pilot pressure from the pilot pump in the lower order, 2nd closing side where bridge pressure acts
It has a pressure receiving surface, and has a closing-side first pressure receiving surface on the top on which an external control pressure controlled by the differential pressure between the maximum load pressure selected by the shuttle valve and the pump pressure acts. It was

Even if this is theoretically aside, in practice, the oil flowing through the throttle of the pressure compensation valve causes turbulence, etc., resulting in flow force, so the pressure difference between the main pump pressure and load pressure and the actuator port discharge flow rate. As for the relationship of, the independence of the pump pressure decreases to the right as indicated by II 'in FIG. As a result, when a plurality of actuators are simultaneously operated, the pressure compensating valve causes hunting and becomes unstable, which is a serious problem.

The present invention was devised to solve the above-mentioned problems, and its purpose is to simplify the passage structure while incorporating a pressure compensating valve and a shuttle valve, and to provide a high degree of freedom in control. Even if the maximum load pressure fluctuates, the pressure compensation controllability is good, and even when multiple actuators are operated simultaneously, the pressure compensation valve does not cause hunting and can be in a stable state. Moreover, it is an object of the present invention to provide a hydraulic control valve that can realize this at a low cost without requiring a major valve modification.

[Means for Solving the Problems]

In order to achieve the above object, the present invention is arranged between a single hydraulic pump and a plurality of actuators driven by the single hydraulic pump, and selects a high-pressure side of the load pressure of the actuator in addition to the directional control valve for the valve body. The valve body has a vertical hole orthogonal to the horizontal hole that slides the spool of the directional control valve, and the upper part of the valve, which incorporates a shuttle valve that sends the signal pressure and a pressure compensation valve that has the function of dividing the discharge oil of the main pump. The pressure compensating valve is arranged in the side vertical hole, the shuttle valve is arranged in the lower vertical hole, and an oil chamber for introducing the load pressure of the actuator is formed at the intersection of the vertical hole and the horizontal hole. Of the balance piston of the pressure compensation valve as the inlet of the
The pressure receiving surface faces, and the balance piston has an open second pressure receiving surface in contact with the pilot pressure from the pilot pump in the vicinity of the open first pressure receiving surface, and the upper side is a closed side on which bridge pressure acts. A first closed side having a second pressure receiving surface, on the top of which an external control pressure controlled by a differential pressure between the maximum load pressure selected by the shuttle valve and the pump pressure acts.
In a hydraulic control valve having a pressure receiving surface, a small diameter hole, an intermediate diameter hole and a large diameter hole are sequentially formed in the vertical hole from the oil chamber upward, while the balance piston has a diameter corresponding to the small diameter hole. A small-diameter land portion, a first land portion having a diameter corresponding to the intermediate diameter hole, and a second land portion having a diameter corresponding to the large-diameter hole are provided, and the opening-side first pressure receiving surface is arranged in the axial direction with the tip of the small-diameter land portion. The second pressure receiving surface on the opening side is formed at the boundary between the small-diameter land portion and the first land portion, and the pump is provided on the rod portion between the first land portion and the second land portion. A through hole for introducing pressure is formed, and an opening-side third pressure receiving surface on which a pump pressure acts is formed at a boundary portion between the rod portion and the second land portion.

〔Example〕

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

1 to 4 and 7 show an embodiment of the hydraulic control valve according to the present invention, FIG. 5 shows an example of use of the hydraulic control valve according to the present invention, and FIG. 6 shows hydraulic pressure according to the present invention. It shows an example of a hydraulic circuit to which a control valve is applied. The hydraulic control valve according to the present invention is
First, as shown in FIG. 1, a plurality of control valves M interposed between a single main pump P and a plurality of actuators S, and an unload relief connected to a main pump line upstream of the control valve M. Valve 600
It has and.

The control valve M includes a directional switching valve 100 for each actuator, a pressure compensating valve 200 with a flow dividing function for compensating the amount of oil flowing through the directional switching valve 100, and a maximum load of the load pressures acting on each actuator. A shuttle valve 300 for selecting a pressure and a valve body 1 are incorporated.

Further, in this hydraulic circuit, a plurality of electromagnetic proportional pressure control valves 800 for creating and supplying external control pressure to each of the pressure compensation valves 200, and a main pump discharge passage downstream of the unload relief valve 600 were connected. A differential pressure detector 810 and a control device 805 connected to the outlet side of the differential pressure detector 810 and controlling the electromagnetic proportional pressure control valve 800 by a signal from this are provided.

The control valve M may be of a type in which the directional switching valve 100, the pressure compensating valve 200 and the shuttle valve 300 are incorporated in a plurality of bodies and stacked, but in this embodiment,
A block type valve body 1 is used, and a plurality of control valves M for each actuator are provided therein. As shown in FIG. 5, this valve body 1 has an unload relief valve 600 and an end plate 6 on both sides.
50 is set and integrated by a tie rod or the like.

The valve body 1 of the control valve M has a first
As shown in FIG. 3 to FIG.
A vertical hole 3 is provided so as to be orthogonal to the horizontal hole 2. The spool 4 of the direction switching valve 100 is slidably inserted into the lateral hole 2, while the vertical hole 3 divided by the spool 4 has a pressure compensating valve 200 on the upper side and a shuttle valve on the lower side. 3
00 is installed respectively. Similar to a known switching valve, both ends of the spool 4 project from the valve body 1 and one end thereof is urged by a return spring mechanism to be returned to the neutral position as shown in FIG. The other end is attached with an operating lever or faces the hydraulic pilot chamber.

An actuator S is provided at the intersection of the horizontal hole 2 and the vertical hole 3.
The oil chamber 20 for introducing the load pressure PL of
Bridge-shaped supply port PA, P
B, actuator ports A and B, and tank ports T and T
Are provided, and the supply ports PA and PB rise up in the valve body and communicate with the vertical hole 3. Tank port T is third
As shown in the figure, it extends at right angles to the plane of the drawing by the common passage.

Rod portions 30 and 30 each having a throttle 31 are formed on the outer periphery of the spool 4 at positions corresponding to the actuator ports A and B, and the ports PA, PB, A, B and T are all blocks at the neutral position shown in the drawing. PA when spool 4 moves to the right
→ A, B → T connection, PA → B, A → T when moving to the left
It is configured in a communication relationship that is a connection.

The spool 4 is not solid but has a connecting passage 32 in the axial direction.
It has A and 32B. These communication passages 32A, 32B are for guiding the load pressure PL of the actuator S to the oil chamber 20, the rear end is closed by the plugs 5a, 5b, the front end is closed at the spool central portion, and the rod portion 30,3
Small holes 34a, 35a, 34b, 35b are provided in the vicinity of 0 and in the region of the oil chamber 20, respectively, which are opened on the outer peripheral surface.

The small holes 34a, 35a, 34b, 35b allow the oil chamber 20 to communicate with the left and right tank ports T, T when the spool 4 is in the neutral position shown in the figure, and when the spool 4 moves, the actuator ports A, B Of these, the load pressure is introduced into the oil chamber 20 from the port used as the supply side. That is, when the spool 4 moves to the right, the small hole 34 on the left side
While the oil chamber 20 and the actuator port A are communicated with each other by a and 35a, the small holes 34b, 35b on the right side block the oil chamber 20 and the actuator port B from each other, and when the spool 4 moves to the left side, the oil chamber is reversed. 20 and the actuator port B are communicated with each other, and the oil chamber 20 and the actuator port A are cut off.

As shown in FIG. 3, the upper vertical hole of the vertical holes 3 is the inner flange-shaped butting wall 12 from the upper surface to the lower bottom of the valve body 1.
And the abutment wall 12 is formed by the through hole 13 in the oil chamber.
It communicates with 20. The upper vertical hole is a small diameter hole 51 having a diameter of D ₃ from the abutting wall 12 to a required height, and the lower area of the small diameter hole 51 communicates with the oil chamber 20 to guide the load pressure PL.
The oil chamber Y ₁ is recessed. A second oil chamber Y ₂ that guides the pilot pump pressure Pi is recessed above the small diameter portion 51, and further, between the second oil chamber Y ₂ and the collecting portion of the supply ports PA and PB. Is formed with a pump pressure chamber Y ₅ for guiding a pump line pressure (hereinafter referred to as pump pressure) Pd, and the second oil chamber Y ₂
The vertical hole between the pump pressure chamber Y ₅ and the pump pressure chamber Y ₅ is an intermediate diameter hole 53 having a diameter D ₂ . A vertical hole between the second oil chamber Y ₂ and the collecting portion of the supply ports PA and PB and above the collecting portion is a large diameter hole 52 having a diameter D ₁ . That is, the three hole diameters are D ₁ > D ₂ > D ₃
It has a relationship of.

The second oil chamber Y _{2 of} each control valve and the pump pressure chamber
As shown in FIG. 3, Y ₅ are connected to each other by common passages 700 and 710 passing through the valve body 1, and are connected to the main pump P and the pilot pump Pi by external piping as shown in FIG. There is.

The pressure compensating valve 200 is shown in FIGS. 1 and 2, and includes a balance piston 6 slidably inserted in a vertical hole, a plug 7 inserted into the upper portion of the balance piston 6, and a balance piston. 6 and a load check valve 8 incorporated in an intermediate portion.

First, the balance piston 6 is provided with a stepped upper hole 60 having a depth from the upper end to a portion corresponding to the pump pressure chamber Y ₅ and a lower end so that a partition wall is provided between the stepped upper hole 60. It is recessed and has a substantially cylindrical shape with the prepared hole 61 that constitutes the opening-side first pressure receiving surface A ₃ . A female thread is provided at the opening of the upper hole 60, and the threaded portion of the plug 7 is screwed onto the female thread to be integrated with the balance piston 6.

Further, the balance piston 6 has an annular tip surface abutting against the abutting wall 12, and the tip region is a small diameter land portion 69 having a diameter d ₃ corresponding to the diameter D ₃ of the small diameter hole 51. Land section
The annular front end surface of 69 and the bottom surface of the prepared hole constitute an opening-side first pressure receiving surface A ₃ that receives the load pressure PL from the oil chamber 20.
A spring 17 is arranged between the bottom surface of the prepared hole and the butting wall 12. The spring 17 is for absorbing vibration, has a small spring force, and the acting force on the balance piston 6 is negligibly small.

The small-diameter land portion 69 ends in the middle of the second oil chamber Y ₂ , and a diameter corresponding to the diameter D _{2 of the} intermediate diameter hole 53 is provided at that portion through a step portion serving as the opening-side second pressure receiving surface A _4. The first land portion 62 of d ₂ is provided, and the pilot pressure is applied to the opening-side second pressure receiving surface A _4.
Pi is working.

The upper end of the first land portion 62 ends near the lower portion of the pump pressure chamber Y ₅ , and from this portion through the rod portion, the second land portion 63 and the second land portion 63 having a diameter d ₁ that matches the diameter D ₁ of the large diameter hole 52, respectively. 3 land part 64
Are tiered. The rod portion between the second land portion 63 and the first land portion 62 faces the pump pressure chamber Y ₅ , and a plurality of through holes 65 for introducing the pump pressure Pd therein are provided in this portion. The stepped portion at the boundary between the portion and the second land portion 63 is the opening-side third pressure receiving surface A ₅ that receives the pump pressure Pd.

On the other hand, the rod portion between the second land portion 63 and the third land portion 64 faces the supply ports PA and PB, and the pressure oil (hereinafter referred to as bridge pressure) Pz that reaches the spool 4 from the supply ports PA and PB at this portion. Is provided with a plurality of small holes 66.

A poppet-type load check valve 8 is slidably fitted in the upper hole above the through hole 65 for introducing the pump pressure Pd therein, and the spring seat plug 14 is screwed to the plug 7.
A spring 80 having a weak spring force is interposed between the valve seat portion and the valve seat portion having an inner flange shape.

The pump pressure oil is supplied to the downstream side of the load check valve 8, that is, the second land portion 63 near the seat portion, from the supply ports PA and PB.
In order to lead to the above, a plurality of supply holes 67 opening in the radial direction are formed as shown in FIG. And supply port
An annular groove (notch) for throttling is provided in the vertical hole at the assembly site of PA and PB so as to obtain an appropriate overlap with the supply hole 67.
22 is formed so that when the balance piston 6 is displaced upward, the control oil amount is guided from the supply hole 67 to the supply ports PA and PB according to the displacement amount.

The plug 7 integrally includes an intermediate collar 70 that is in close contact with the upper end surface of the balance piston 6 following a portion screwed into the female screw of the balance piston 6, and a head 71 having a diameter d ₄ smaller than the small diameter hole 51. The head 71 is slidably inserted into the boss 90 of the cap assembly 9 fitted in the vertical hole.
A third hydraulic pressure Y ₃ is formed at the lower end of 90, the intermediate collar 70, and a portion surrounded by the vertical hole, with the intermediate collar 70 serving as the closing-side second pressure receiving surface A ₂ . The boss 90 is sealed with a vertical hole by an O-ring and is held by a connector 91 having a port C for introducing an external control pressure. In the region surrounded by the connector 91, the boss inner surface and the head end surface, A fourth oil chamber Y ₄ is formed having the head end surface as the first pressure receiving surface A ₁ on the closing side. The connector 91 is fixed to the valve body 1 by an appropriate method.

Back pressure chamber 81 of the load check valve 8, that is, a spring
The chamber accommodating 80 communicates with the small hole 66, and its back pressure chamber 81 guides the bridge pressure Pz, so that the axial hole 140 and the lateral hole 141 formed from the spring seat plug 14 to the head 71. Communicates with the third oil chamber Y ₃ .

The relationship between the pressure receiving surface of the balance piston 6 and the hydraulic pressure acting on it is shown schematically in FIG.

That is, first, the external control pressure Pc (pressure set according to the pressure difference between the maximum load pressure Pl and the pump pressure Pd) introduced from the port C acts on the closing-side first pressure receiving surface A ₁ . Closed side second
The pressure receiving surface A ₂ is a ring-shaped area of the intermediate collar 70, and the bridge pressure Pz introduced in the route of the small hole 66-vertical hole 140-lateral hole 141 works here.

In contrast, the first pressure receiving surface A ₃ side open load pressure PL of the actuator is introduced through the oil chamber 20 from the communication passage 32A or 32B by the movement of the spool 4 acts. In FIG. 7, the shape of the opening-side first pressure receiving surface A ₃ is shown as a solid shape for the sake of clarity. Further, the opening-side second pressure receiving surface A ₄ is a ring-shaped area formed by the difference between the first land portion 62 and the small diameter land portion 69, and is a portion where the pilot pressure Pi acts. Finally, the third pressure receiving surface A ₅ on the opening side is provided with the first land portion 62 and the second land portion.
It is the ring-shaped area consisting of the difference with 63,
Pd works.

That is, in the present invention, (A ₅ × Pd) in the opening direction of the balance piston _{6 + (A 4 × Pi)} + (A 3 + PL) ... force acts in (1), closing in the direction (A ₂ × Pz) + (A ₁ × Pc) (2) acts, and by the balance between the resultant force of the three forces in the opening direction and the resultant force of the two forces in the closing direction, the supply hole 67 and the annular groove 22.
The opening degree of the diaphragm consisting of and is controlled.

Now, assuming that A ₁ = A ₄ , A ₃ = A ₂ −A ₅ , A ₁ + A ₂ = A ₅ + A ₄ + A _3, and substituting this into the formulas (1) and (2), the balance piston 6 becomes The following balance relationship is established.

(A ₂ × Pz) + (A ₁ × Pc) = (A ₅ × Pd) + (A ₁ × Pi) + (A ₂ −A ₅ ) ・ PL (A ₂ × Pz) − (A ₂ −A ₅ ) ・ PL = (A ₅ × Pd) + (A ₁ × Pi) − (A ₁ × Pc) Therefore, This is a feature of the present invention, which means that the independence of the pressure compensating valve with respect to the pump pressure Pd tends to rise to the right as shown by the trend lines I and I'in FIG.

Next, the shuttle valve 300 is shown in FIGS. The shuttle valve 300 includes a holder 301 that is oil-tightly inserted while being positioned by a non-circular or eccentric flange 301a in a lower vertical hole, a cap 302 screwed to the tip of the holder 301, and a cap 302. A ball valve 303 housed in the valve body housing hole 301b between the holders 301 and a plug 305 for fixing the holder 301 are provided.

The ball valve 303 has a tip of a cap 302 and a valve body accommodation hole 301b.
The cap 302 has a first inlet hole 302a for introducing a load pressure of the actuator S to which the control valve belongs from the oil chamber 20 into the valve body accommodating hole 301b. Has been drilled.

On the other hand, on the outer periphery of the holder 301, as shown in FIG. 4, recesses 301e and 301f are provided which do not communicate with each other due to a displacement of 180 degrees, and one recess 301e is provided at the bottom of the valve body accommodation hole 301b. 2nd by communicating through the drill hole 301c and the communication hole 301g
An inlet hole is formed, and the other concave portion 301f communicates with the valve body accommodation hole 301b through the communication hole 301h to form an outlet hole. Further, the valve body 1 is provided with passages 15 and 16 communicating with the recess 301e and the recess 301f so as to be orthogonal to the vertical holes.

In the valve body accommodation hole 301b of the shuttle valve 300, the load pressure of the actuator is passed through the oil chamber 20 from the first inlet hole 302a, and the load pressure from the adjacent actuator is passed through the passage 15.
If the load pressure on the side of the second inlet hole is high, the ball valve 303 closes the seat on the side of the cap, and if the load pressure on the side of the first inlet hole 302a is high, the valve element housing hole bottom is introduced. The seat on the side is closed and the next shuttle valve is reached from the communication hole 301h through the passage 16. And here too, a similar selection is made,
The maximum load pressure Pl is taken out from the final shuttle valve. The maximum load pressure Pl is guided to the passage 18 from the valve body of the control valve M at the right end as shown in FIG. 6, and is sent to the differential pressure detector 810 and the unload relief valve 600 by the branch passages 180 and 181. .

Details of the unload relief valve 600 are shown in FIG. This unload relief valve 600 has an unload valve 600A on the right side area of the body 601 and a relief valve 600B on the left side area.
Are arranged. Needless to say, the unload valve 600A essentially releases the pressure oil discharged from the main pump P at a low pressure when the directional control valve is not operated, and the relief valve 600B reaches the set pressure. When this happens, the entire amount of pressure oil from the main pump is released to the tank.

In detail, a pump passage 604 and tank passages 605 and 615 are formed on both sides of the body 601, and the pump passage 604 and the tank passage 60 are formed.
One end of the valve 5 opens to the mating surface with the control valve, and the other end opens to the pump port and tank port of the centralized piping surface (not shown).

A bush 612 is inserted and fixed in a valve hole orthogonal to the pump passage 604 and the tank passage 605.
A tip of a plug 603 screwed into the 601 from the open side is inserted, and an unload valve body 602 guided by a bush 612 is slidably inserted in a deep part of the valve hole.

The unloading valve body 602 has two coaxial blind holes 606 and 610 formed from both ends, and a spring 611 is arranged between the bottom of the left blind hole 610 and the tip of the plug 603. As a result, the unloading valve body 602 is always urged to the right side, thereby forming a load pressure chamber (back pressure chamber). Hereinafter, the blind hole 610 is referred to as a load pressure chamber. A passage hole 620 is formed in the middle of the unload valve body 602 to connect the pump passage 604 and the right blind hole 606, and a pressure receiving chamber (pilot chamber) 607 is formed at the mouth end of the blind hole 606. Has been done.

On the other hand, a spring chamber 613 is formed in the plug 603,
A passage hole 614 which is always in communication with the tank passage 615 is provided on the side of the tip end thereof, and in the axial direction of the spring chamber 613,
A passage hole for connecting the load pressure chamber 610 and the tank passage 615.
616 has been drilled. A pilot type relief valve element 617 for opening and closing the passage hole 616 is arranged in the spring chamber 613, and a spring 619 arranged between the adjusting screws 618 at the rear end.
Is always biased toward the closing side.

The bush 612 is formed with a throttle 609 communicating with the load pressure chamber 610, and the throttle 609 communicates with a signal passage 608 bored from the mating surface of the body 601.

The mating surfaces of the unload relief valve 600 and the control valve M configured as described above are brought into close contact with each other, and the pump passage 604 and the pump port Y ₅ , the tank passages 605 and 615 and the tank port T, and the pilot pump passage (not shown). ) And a common passage 710 as a pilot pump port, and the signal passage 608 communicates with the final shuttle valve 300 outlet (branch passage 181).

The main pump P is connected to the pump passage 604 of the unload relief valve 600, the pilot pump Pi is connected to the pilot pump passage, and the tank passages 605 and 615 are connected to the tank. A relief valve 700 is connected to the pilot line 19 from the pilot pump Pi as shown in FIG. 6 to keep the pilot pump pressure Pi constant. Further, the pilot line 19 is connected to the inlet side of each three-port two-position switching type electromagnetic proportional pressure control valve 800 provided for each actuator, and the outlet side of the electromagnetic proportional pressure control valve 800 is each pressure compensation valve 200. The external control pressure Pc is applied to the first pressure receiving surface A ₁ on the closing side via the fourth oil chamber Y ₄ .

A control unit 805 that individually sends a control signal (current) is connected to the electromagnetic unit side that moves the spool of each electromagnetic proportional pressure control valve 800 against the spring. The controller 805 is connected to the signal outlet of the differential pressure detector 810. The differential pressure detector 810 is interposed in the discharge passage of the main pump P and the maximum load pressure delivery passage from the final shuttle valve 300 as described above, and the differential pressure between the pump Pd and the maximum load pressure Pl (Pd-
Pl) is detected, and its magnitude is converted into a current value and output. The control device 805 calculates a control value based on the current value from the differential pressure detector 810. That is, the differential pressure detector
When the output from 810 is large (when Pd-Pl is large), a lower signal current value is sent to the solenoid proportional pressure control valve 800, and as the differential pressure is smaller, a higher signal current value is transmitted to the solenoid proportional pressure control valve 80.
Send to 0.

As a result, the electromagnetic proportional pressure control valve 800 is controlled by the external control pressure Pc.
= Pi− (Pd−Pl) is sent to the first pressure receiving surface A ₁ on the closing side, and in the pressure compensating valve 200, the pilot pump pressure Pi and the external control pressure Pc
It is controlled so that the differential pressure between and is equal to the differential pressure between the main pump pressure Pd and the maximum load pressure Pl.

The controller 805 also has a function of individually setting the output to each electromagnetic proportional pressure control valve 800. As a result, the output to a certain electromagnetic proportional pressure control valve 800 is increased or decreased, and the pressure difference between the pressure Pi of the second oil chamber Y ₂ and the fourth oil chamber Y ₄ is increased or decreased, thereby opening the throttle 67. The function of the pressure compensating valve 200 can be changed by adjusting the degree to enable the combined operation. If necessary, the output to a certain electromagnetic proportional pressure control valve 800 is set to zero (external control pressure Pc is set to zero), and the difference between the pressure Pi in the second oil chamber Y ₂ and the fourth oil chamber Y ₄ is set. The function of the pressure compensation valve 200 can be released by setting the pressure to the maximum and opening the throttle 67 fully.

It should be noted that the present invention does not preclude the incorporation of other valves in the valve body. That is, for example, a passage in which an overload relief valve is inserted between the tank ports T, T and the actuator parts A, B, or a part of the actuator port B or A communicates with the actuator port B or A and the other end communicates with the tank port. Is provided, and a non-leak valve for locking is attached to this.

[Operation of Example]

 Next, the operation of the hydraulic control valve device according to the present invention will be described.

The pressure oil discharged from the main pump P enters the pump passage 604 of the unload relief valve 600. When each directional control valve 100 is in the neutral position, the oil chamber (load pressure introducing port) 20 is connected to the tank port T by the connecting passages 32A and 32B as shown in FIG.
Because it communicates with the oil chamber of all control valves
The pressure of 20 and the pressure selected by the shuttle valve 300 are both low, and the maximum load pressure Pl input to the signal passage 608 of the unload relief valve 600 is also low, so the load pressure chamber 610 is kept at low pressure. Be drunk Therefore, pilot room 6
Pump pressure Pd of 07 resists spring 611 and unloads valve 6
In order to move 02 to the left in FIG. 8, the pump passage 604 and the tank passage 605 communicate with each other, and the discharge oil of the main pump is returned to the tank with no load.

When the spool 4 of the directional control valve 100 of any one of the control valves M is moved from the neutral position, the load pressure from the actuator is applied to the oil chamber 20 via the communication passage 32A or 32B.
PL is introduced, and this pressure is applied to shuttle valve 300 and signal passage 608.
Into the load pressure chamber 610 of the unload valve. As a result, the unload valve element 602 moves to the right as shown in FIG. 5 and closes the unload valve 600A.

When the maximum load pressure Pl selected by the shuttle valve 300 and introduced into the six load chambers 610 reaches a certain pressure set by the adjusting screw 618, the relief valve body 617 resists the spring force of the spring 619 and moves to the left. Move to. As a result, the pressure in the load pressure chamber 610 is lowered and a differential pressure is generated in the unload valve body 602, so that the valve body moves to the left and the pressure oil in the pump passage 604 escapes to the tank passage 605. Now, when the spool 4 of the directional control valve 100 incorporated in the control valve M is moved, the pressure oil supplied from the common passage 700 to the pump pressure chamber 70 passes from the pressure compensating valve 200 through the directional control valve 100 to the actuator S. Flow to.

That is, when the spool 4 is moved to the right, the pressure oil in the pump pressure chamber 70 enters the upper hole 60 from the through hole 65 of the balance piston 6, opens the load check valve 8 against the spring 80, and supplies it. Supply port PA, PB from annular groove 22 through hole 67
The flow rate is controlled by the throttle 31 of the spool 4 and then supplied to the actuator, for example, the head side of the cylinder via the actuator port A. At the same time, the oil on the rod side of the actuator moves to the actuator port B.
Then, it is returned to the tank through the throttle 31 and the tank port T.
When the spool 4 is moved to the left, the pressure oil reaches the rod side of the actuator along the route of the supply port PB → throttle 31 → actuator port B, and the oil on the head side is the actuator port A → throttle 31 → tank port. It is returned to the tank by the route of T.

On the other hand, the pilot pump Pi is driven at the same time as the main pump P, the pilot pressure Pi controlled to a constant pressure by the relief valve 700 reaches the valve body 1 from the passage 19, and the common passage 71
From 0 to the second oil chamber Y ₂ , the opening side second pressure receiving surface A ₄ located there is pressed, and branched from the passage 19 and sent to the inlet side of each electromagnetic proportional pressure control valve 800. .

When the spool 4 moves to the right as described above, the small hole 35b of the right communication passage 32B is closed by the inner wall of the lateral hole 2, and the small hole 34a of the left communication passage 32A communicates with the actuator port A. On the contrary, if the spool 4 moves to the left, the small hole 34b of the communication passage 32B on the right side communicates with the actuator port B.

As a result, the load pressure PL from the actuator is introduced into the oil chamber 20. Then, the load pressure PL of the oil chamber 20 applies an opening side force to the balance piston 6 by the opening side first pressure receiving surface A ₃ of the first oil chamber Y ₁ , while the shuttle valve 300 passes through the first inlet hole 302a. Flow into.

Further, the pump pressure Pd acts as an upward force on the opening-side third pressure receiving surface A ₅ located in the pump pressure chamber Y ₅ .

Although the balance piston 6 rises due to the three pressures on the opening side, the pump discharge oil flows through the supply hole 67, flows from the annular groove 22 to the supply ports PA and PB, and its pressure (bridge pressure) Pz.
From the radial small hole 66 into the back pressure chamber 81 of the load check valve 400, flow into the third oil chamber Y ₃ through the shaft hole 140 and the lateral hole 141, and reach the closing side second pressure receiving surface A ₂ on the balance piston. Serves as closing side pressure of 6.

A load pressure is introduced into the shuttle valve 300 from another adjacent shuttle valve 300 through the second inlet hole, and the ball valve 303 moves depending on the level of the pressure, so that the higher load pressure is passed through the passages 16,15.
After reaching the next shuttle valve 300, the maximum load pressure Pl is taken out from the last shuttle valve, sent to the differential pressure detector 810, and sent to the unload relief valve 600 as the closing side pilot pressure. In the differential pressure detector 810, the main pump pressure Pd and the maximum load pressure Pl are compared, and a current according to the differential pressure is sent to the control device 805, where the control current is calculated and the electromagnetic proportional pressure control valve 800 operates. External control pressure P
c is made. The external control pressure Pc is Pc = Pi− (Pd−P
l) That is, the pressure is set according to the pressure difference between the maximum load pressure Pl and the pump pressure Pd. And this external control pressure
Pc is from the port C of the cap assembly 9 to the fourth oil chamber Y ₄
Is introduced into the balance piston 6 and acts as the closing side pressure of the balance piston 6 by the closing side first pressure receiving surface A ₁ .

In the pressure compensation valve 200, when the balance piston 6 is displaced upward, the throttle mechanism including the annular groove 22 and the supply hole 67 is opened, and when the balance piston 6 is displaced downward, the throttle is closed. And the first
The load pressure PL of the actuator S is introduced into the oil chamber Y ₁
The pilot pump pressure Pi is introduced into the oil chamber Y _2, and the resultant force acts as a force for opening the throttle.

On the other hand, the bridge pressure Pz is introduced into the third oil chamber Y ₃ , the external control pressure Pc described above is introduced into the fourth oil chamber Y _4, and the resultant force acts as a pressure to close the throttle. The throttle opening of the pressure compensation valve 200 can be controlled by the balance of the resultant force of the two forces and the resultant force of the two closing directions.

More specifically, movement of the spool 4 of the directional control valve 100 causes a throttle corresponding to one actuator port A or B.
When the opening degree of 31 increases, the load pressure PL increases, so that the throttle opening degree of the pressure compensation valve 200 increases, which increases the flow rate of the throttle 31 and increases the amount of oil supplied to the actuator. Conversely, if the opening of the throttle 31 of the directional control valve 100 is reduced,
The throttle opening of the pressure compensation valve 200 is reduced, and the amount of oil supplied to the actuator is reduced. Therefore, the directional control valve 100
It is possible to control the amount of oil supplied to the actuator, that is, the drive speed of the actuator, in accordance with the operation amount of.

When the load of the corresponding actuator increases and the load pressure PL increases, the throttle opening of the pressure compensation valve 200 increases so as to increase the bridge pressure Pz, and in the opposite case, the bridge pressure Pz increases.
Since the throttle opening is decreased so as to lower z, the oil supply amount per unit time to the actuator can be maintained according to the operation amount of the directional control valve 100, regardless of the load change of the actuator.

The present invention uses the external control pressure Pc, and the larger the external control pressure Pc is, that is, the larger the maximum load pressure Pl is, the more the flow rate of the entire circuit is throttled. The total amount of oil required by the actuator is limited according to the size of the pump, the shortage of the amount of oil discharged from the pump is alleviated, and the light load actuator and the heavy load actuator are controlled in a well-balanced manner.

Therefore, the pressure-receiving surface on the opening side of the balance piston 6 does not have three stages as in the present invention, and FIG. 8 (corresponding to the third prior art).
As shown in, the large-diameter hole of the small diameter hole 51 and the diameter D ₁ of the diameter D ₃ of the vertical hole 3
52, and the diameters of the first land portion, the second land portion, and the third land portion of the balance piston 6 are all the same diameter d ₁ , the pump pressure Pd of the pump pressure chamber Y ₅ is equal to that of the land portion. Since the pressure receiving areas are equal, they are offset at the top and bottom, and the pump pressure Pd does not work as a force on the opening side and the closing side of the balance piston 6.
That is, the pressure receiving surfaces in this case are the opening side first pressure receiving surface A ₃ and the opening side second pressure receiving surface A ₄ : the closing side first
There is a relationship between the pressure receiving surface A ₁ and the second pressure receiving surface A ₂ on the closing side.

Under the conditions of A ₁ = A ₄ and A ₂ = A ₃ , this is (A ₁ × Pc) + (A ₂ × Pz) = (A ₃ × PL) + (A ₄ × Pi) , A ₂ (Pz−PL) = A ₁ (Pi−Pc), and Pz−PL = A ₁ / A ₂ (Pi−Pc). That is, the differential pressure between the bridge pressure and the load pressure becomes the differential pressure between the pilot pressure and the external control pressure, and
Proportional according to the ratio of the pressure receiving surface. This is theoretically valid. However, when it is actually operated, the actuator port flow rate tends to decrease even though the actuator port flow rate Q should originally be constant even if there is a pressure difference between the pump pressure Pd and the load pressure PL. It becomes more remarkable as the differential pressure of Pd-PL increases.

That is, even if it is designed like II in FIG.
It falls to the right like the actual line of I '. This is because the oil flowing through the throttle of the pressure compensating valve 200 causes a turbulent flow or the like to generate a flow force, and the degree of downward sloping is various.

If the independence of pump pressure tends to decrease to the right, the system becomes unstable and hunting occurs.
That is, if a plurality of actuators are operated simultaneously and there is an actuator with a large load, the pump pressure Pd rises to match it.

As the pump pressure Pd increases, the throttle opening decreases, and the actuator port flow rate Q decreases.

When the actuator port flow rate Q decreases, the load sensing differential pressure ΔPLs (= Pd-Pl) increases. That is, there is a margin in the pump discharge amount.

When the load sensing differential pressure ΔPLs thus rises, the differential pressure ΔPc between the pilot pressure Pi and the external control pressure Pc rises.

When the differential pressure ΔPc increases, the balance piston 6 moves in the opening direction in an attempt to increase the actuator port flow rate Q.

When the balance piston 6 opens, the pump pressure Pd drops.

When the pump pressure Pd decreases, the actuator port flow rate Q increases and becomes larger than the initial value. And
As the actuator port flow rate Q increases, the load sensing differential pressure ΔPLs decreases.

As a result, the opposite phenomenon to the above occurs, and as a result, the above phenomena are repeated.

On the other hand, in the present invention, the balance piston 6
Is the diameter of the small land portion corresponding to the diameter D ₃ of the small diameter hole 51 and the diameter of the first land portion 62 and the diameter of the large diameter hole 52 corresponding to the diameter D _{2 of the} intermediate diameter hole.
Has a second land portion 63 having a diameter corresponding to D ₁ , and the first land portion
An opening-side third pressure receiving surface A ₅ on which the pump pressure Pd acts is formed at a boundary portion between the rod portion between the 62 and the second land portion 63 and the second land portion 63. That is, as the open side pressure receiving surface, the load pressure
1st pressure receiving surface A ₃ which receives PL and 2nd which receives pilot pressure Pi
In addition to the pressure receiving surface A ₄ , a third pressure receiving surface A ₅ that receives the pump pressure Pd is formed.

As a result, the balance between the open side and the closed side is A ₃ + A ₄ + A ₅ = A ₁ +
It becomes A ₂ , and the independence of the third pressure receiving surface A ₅ with respect to the pump pressure Pd tends to rise to the right beforehand (I, I ′ in FIG. 10).
Therefore, although the total pressure receiving area is the same as that shown in FIG. 8, the actuator port flow rate Q does not decrease even if the pump pressure Pd rises, and the pump pressure Pd and the load pressure PL
If the difference between the two increases, the actuator port flow rate Q increases accordingly. Therefore, the above-mentioned hunting does not occur, an operation destabilization phenomenon caused thereby can be appropriately avoided, and a stable control system state can be obtained.

〔The invention's effect〕

According to the present invention described above, since the pressure compensating valve and the shuttle valve are housed in the vertical hole intersecting with the lateral hole for housing the spool while having the switching valve, the pressure compensating valve, and the shuttle valve, the structure can be made compact. As a mechanism for introducing the load pressure of the actuator to the pressure compensating valve and the shuttle valve, an oil chamber is formed at the intersection of the vertical hole and the lateral hole, and the pressure compensating valve and the shuttle valve are faced here. The load pressure introducing passage structure can be simplified, and the external control pressure set according to the differential pressure between the maximum load pressure and the pump pressure is used as the one-side pressure that applies the closing-side force to the pressure compensation valve. Therefore, even if the maximum load pressure fluctuates, the degree of freedom of controllability of each actuator can be improved.

Moreover, the vertical hole 3 is provided with a small diameter hole 51, an intermediate diameter hole 53 and a large diameter hole 52 in order upward from the oil chamber 20, while the balance piston 6 has an intermediate portion with a small diameter land portion 69 having a diameter corresponding to the small diameter hole 51. A first land portion 62 having a diameter corresponding to the diameter hole 53 and a second land portion 63 having a diameter corresponding to the large diameter hole 52 are provided, and the opening side first pressure receiving surface A ₃
Is constituted by the tip of the small-diameter land portion 69 and the prepared hole 61 snarling in the axial direction, and the opening-side second pressure receiving surface A ₄ is formed at the boundary portion between the small-diameter land portion 69 and the first land portion 62, and First land part
A through hole 65 for introducing pump pressure is formed in the rod portion between the 62 and the second land portion 63, and the third pressure on the opening side where the pump pressure Pd acts on the boundary portion between the rod portion and the second land portion 63. Since the surface A ₅ is formed, the independence of the opening-side third pressure receiving surface A ₅ with respect to the pump pressure Pd can be increased in advance to the right, whereby hunting can be prevented. Therefore, even if the total pressure receiving area is the same as the two-stage pressure receiving surface, the actuator port flow rate Q does not decrease due to the increase of the pump pressure Pd, and if the difference between the pump pressure Pd and the load pressure PL becomes large, the corresponding Appropriate stability in which the actuator port flow rate Q increases can be provided. Moreover, the excellent effect that such characteristics can be realized at a low cost without requiring a large change in the valve body or the balance piston is obtained.

[Brief description of drawings]

1 is a sectional view showing an embodiment of a hydraulic control valve according to the present invention, FIG. 2 is a partially enlarged view thereof, FIG. 3 is a sectional view of a valve body of a control valve, and FIG. 4 is a shuttle according to the present invention. FIG. 5 is a cross-sectional view showing the mutual connection relationship of the valves, FIG. 5 is an explanatory view showing the cross-section of the unload relief valve and the control valve in the present invention, and FIG. 6 shows an example of a hydraulic control system utilizing the present invention. Circuit diagram, Figure 7 shows the first
FIG. 8 is an explanatory view schematically showing the pressure receiving surface relationship of the embodiment, FIG. 8 is an explanatory view showing a conventional pressure receiving surface, and FIG. 9 is a trend line graph showing independence of the present invention and conventional pump line pressure. P ... Main pump, M ... Control valve, S ...
... actuator, 100 ... directional valve, 200 ... pressure compensation valve, 300 ... shuttle valve, 1 ... valve body, 2 ...
… Horizontal hole, 3 …… Vertical hole, 4 …… Spool, 6 …… Parance piston, 20 …… Oil chamber, 51 …… Small hole, 52 …… Large hole 5
2, 53 …… Medium diameter hole, 61 …… Pilot hole, 62 …… First land part, 63 …… Second land part, 69 …… Small diameter land part, A ₁ ……
Closing side first pressure receiving surface, A ₂ ... Closing side second pressure receiving surface, A ₃ ... Opening side first pressure receiving surface, A ₄ ... Opening side second pressure receiving surface, A ₅ ... Opening side third pressure receiving surface

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Yusuke Kajita 650 Jinrachicho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura factory (72) Genroku Sugiyama 650 Kintate-cho, Tsuchiura, Ibaraki Hitachi Construction Machinery Stock company Tsuchiura factory (56) Reference JP-A-2-134402 (JP, A) JP-A-2-66302 (JP, A) JP-A-1-266301 (JP, A)

Claims (1)

[Claims] 1. A single hydraulic pump is provided between a plurality of actuators driven by the hydraulic pump, and in addition to a directional control valve, a high pressure side of the load pressure of the actuator is selected as a valve body to provide a signal pressure. It incorporates a shuttle valve for sending and a pressure compensating valve that has a function of diverting the discharge oil of the main pump, and the valve body has a vertical hole orthogonal to the horizontal hole that slides the spool of the directional control valve, and the pressure on the upper vertical hole A compensating valve is arranged, a shuttle valve is arranged in the lower vertical hole, an oil chamber for introducing the load pressure of the actuator is formed at the intersection of the vertical hole and the horizontal hole, and the inlet of the shuttle valve faces the oil chamber. , Open side of balance piston of pressure compensation valve
The pressure receiving surface faces, and the balance piston has an open second pressure receiving surface in contact with the pilot pressure from the pilot pump in the vicinity of the open first pressure receiving surface. A first closed side having a second pressure receiving surface, on the top of which an external control pressure controlled by a differential pressure between the maximum load pressure selected by the shuttle valve and the pump pressure acts.
In a hydraulic control valve having a pressure receiving surface, a small diameter hole 51, an intermediate diameter hole 53, and a large diameter hole 52 are sequentially formed in the vertical hole 3 upward from the oil chamber 20, while the balance piston 6 has the small diameter hole. A small-diameter land portion 69 having a diameter corresponding to the hole 51, a first land portion 62 having a diameter corresponding to the intermediate diameter hole 53, and a second land portion 63 having a diameter corresponding to the large-diameter hole 52 are provided. Pressure receiving surface A ₃
Is constituted by the tip of the small-diameter land portion 69 and the prepared hole 61 snarling in the axial direction, and the opening-side second pressure receiving surface A ₄ is formed at the boundary portion between the small-diameter land portion 69 and the first land portion 62, and First land part
A through hole 65 for introducing pump pressure is formed in the rod portion between the 62 and the second land portion 63, and the third pressure on the opening side where the pump pressure Pd acts on the boundary portion between the rod portion and the second land portion 63. A hydraulic control valve characterized in that it forms a surface A ₅ .