JPS5854206A - fluid control device - Google Patents

Abstract

PURPOSE:To enable preventing occurrence of scooping phenomena in such a way that a means of reducing shaft thrust is provided on the fluid path on the return side of a throttle change-over valve, and movement of a spool is stablilized. CONSTITUTION:Shaft thrust W1 due to fluid which is flowed from a pump port P along a path on the feeding side to a load port (A), and shaft thrust W2 due to fluid which is flowed from a load port B along a fluid path on the return side to a reservoir port T are given to a spool 23 of a main valve in a throttle change-over valve. Here, phi2, angle of deflection made by fluid velocity vector which is flowed through a penetrating hole 32 becomes almost 90 deg., since the penetrating hole 32 is made in the orthogonal direction to the axis of the spool 23. In consequence, shaft thrust becomes such as W2=rhoQ2V2cosphi2 0, the spool 23 is stabilized in quiescence, so that scooping will not be allowed to be produced in the hydraulic cylinder 4.

Description

Detailed Description of the Invention The present invention is suitable for use in a mold opening/closing device of an injection molding machine, for example, and is a fluid that can drive an actuator that drives a load with large inertia while smoothly controlling the speed without jerking. Regarding a control device.

The fluid acting on the discharge amount control section of the variable pump is controlled by a load sensing valve that responds to the differential pressure before and after the flow rate adjustment section of the throttle switching valve, and the discharge amount of the variable pump is controlled by the difference between before and after the flow rate adjustment section of the throttle switching valve. In a fluid control device that saves energy by controlling the pressure to be constant and not discharging unnecessary fluid, the throttle switching valve is used to drive a load with large inertia, such as the mold of an injection molding machine. When controlling the speed of an actuator, the system consisting of the throttle switching valve, load sensing valve, and variable pump resonates via the control fluid due to the influence of large inertia, and the actuator connected to the load resonates through the control fluid. It is a common experience that jerking occurs in the actuator.

When this jerking occurs, the device itself is damaged and its commercial value is lost, so the following devices have been proposed in the past. That is, while fluid is supplied from the throttle switching valve to the actuator that drives a large load, the return fluid from the actuator is passed through the throttle switching valve and directly returned to the tank, and the return fluid is supplied to the throttle switching valve. At the same time, a small throttle is installed in the return line between the actuator and the tank to apply back pressure to the actuator, thereby preventing the actuator from overrun (speeding up faster than the steady speed). A fluid control device has been proposed that prevents the occurrence of jerking in the actuator by preventing this.

However, in this fluid control device, since the return side fluid from the actuator does not pass through the throttle switching valve, the direction of the actuator cannot be controlled by the throttle switching valve, and the overall circuit configuration inevitably becomes complicated. Furthermore, since the return line is provided with a diaphragm, the resistance caused by the diaphragm becomes extremely large when the actuator is driven at high speed, making it unsuitable for modern demands for energy conservation.

The purpose of this invention is to eliminate the above drawbacks,
With a simple circuit configuration and without energy loss,
An object of the present invention is to provide a novel fluid control device capable of smoothly driving an actuator without jerking even against a load with large inertia.

In view of this, the present invention provides a variable pump, an actuator, a throttle switching valve having at least four boats connected between the variable pump and the actuator, and a fluid acting on a discharge amount control section of the variable pump. A load sensing valve is provided, and the load sensing valve is made to respond to the differential pressure before and after the flow rate adjusting section of the throttle switching valve to control the discharge amount of the variable pump to control the differential pressure before and after the flow rate adjusting section. is controlled to a constant value, and at least one return side fluid passage of the throttle switching valve is provided with means for reducing the axial thrust of the fluid against the spool, thereby stabilizing the operation of the spool. .

Hereinafter, the present invention will be described in detail with reference to the illustrated embodiments.

In FIG. 1, l is an electromagnetic proportional throttle switching valve consisting of a main valve 2 and a pilot valve 8, 4 is a hydraulic cylinder that drives a load 5 with large inertia, 6 is a variable pump, and 7 is the discharge amount of the variable pump 6. This is a power matching circuit that controls the

The main valve 2 is a closed center type β position switching valve, and has a cylinder chamber 11 formed in the main body lO.
Annular grooves 12.18°, 14.15.16 are provided sequentially from the left. The annular groove 12 is connected to the tank port T, the annular groove 18 is connected to the load port B, and the annular groove 14 is connected to the pump port P.
The annular groove 15 communicates with the load bow A, and the annular groove 16 communicates with the tank bow T. The inner wall 11b between the annular grooves 13 and 14 of the cylinder chamber ↓1 has an opening 18b of the feedback passage 17, and the annular groove 1
An opening tSa of the feedback passage 17 is provided in the inner wall 11a between the holes 4 and 15.

A spool 23 having three lands 20, 21° 22 is slidably fitted into the cylinder chamber 11, and depending on the position of the spool 23, the pump port P is connected to the load port A or B, and the tank port 'r is connected to the cylinder chamber 11. (load baud) B or A.

Further, pilot chambers 2 are provided at both ends of the spool 23, respectively.
5 and 26, and bias coil springs 27 and 28 are fitted into the pilot chambers 25 and 26, respectively.

The land 20 is provided with means 30a for reducing the axial thrust of the fluid against the spool 23. Said means 8
0a consists of an annular groove 31 provided on the right end surface 40 of the land 20 in FIG. The through hole 32 of this means Boa therefore prevents the fluid flowing into the annular groove 12 in the return fluid passage from the load port B to the tank port T when the spool 23 is in the position shown in FIG. is guided through the through hole 32,
It flows in the radial direction, that is, in the direction perpendicular to the axis of the spool 23.

On the other hand, the pilot valve 8 is an electromagnetic pressure reducing valve having a well-known switching function, and includes solenoids 41a and 41b, a primary port m, and secondary ports n, a, and nB. While reducing the fluid pressure of the solenoid 41a to a pressure proportional to the current value supplied to the solenoid 41b, the secondary port nB or n on the side where the pressure reduction is not controlled is controlled.
A is connected to the tank 44.

The primary port m of the pilot valve 3 is connected to the pump port P of the main valve 2 via a line 43 and the annular groove 14, while the secondary port FnB is connected to the pilot chamber 25 of the main valve 2 via a line 45. , t and secondary ports nA are connected to the pilot room 26 via a line 46.

Therefore, when current is applied to the solenoid 41a or 41b of the pilot valve 3, the pressure of the secondary port nA or IB is transmitted to the pilot chamber 26 or 25 of the main valve 2, and the spool 23 is The pressure transmitted to one end surface is displaced to a position where the spring force of the other coil spring 27 or 28 is balanced, and the opening degree of the flow rate adjustment section 70 of the main valve 2 is determined in proportion to the above-mentioned current value. It has become.

On the other hand, the load port A of the main valve 2 is connected to the rod side port R of the hydraulic cylinder 4 via a line 51, and
The load port B is connected to the head side port H of the hydraulic cylinder 4. A pump port P of the main valve 2 is connected to a discharge port of the variable pump 6 via a pump line 53. Tank ports T, T of the main valve 2 are connected to tanks 56, 56, respectively.

On the other hand, the power matching circuit 7 includes a 3-port pilot valve 101 and a pilot relief valve 102 as an example of a load sensing valve.

The load sensing valve lot connects ports A and n with the symbol V, and closes port m, while
At the symbol position v2, port m and port n are communicated and port t is closed, and the spring force of the spring 105 of the spring chamber 104 is, for example, %/
cd, and when the differential pressure between the pilot chamber 106 and the spring chamber 104 is 6 kg/- or more, it is located at the symbol position ■, and the differential pressure is 6 kII/+! In the following cases, the symbol is positioned at symbol position 2.

A pump line 53 is connected to the port t of the load sensing valve 101 via a pilot line 111, and a port 113 is connected to the port m via a pilot line 112. Above load sensing valve 1
A discharge amount control section 115 consisting of a swash plate control cylinder which is always biased toward the maximum discharge side by, for example, a spring of the variable pump 6 is connected to port n of 01 via a pilot line 114. Further, a pump line 58 is connected to the pilot chamber 106 of the load sensing valve 101 via a pilot line 116, and the spring chamber 104 is connected to the feedback passage 17 of the main valve 2 via a pilot line 119 having a throttle 118. are doing. A tank 121 is connected between the spring chamber 104 of the load sensing valve 101 and the throttle 118 via a pilot line 120 provided with an intermediate pilot relief valve 102.

Therefore, the pilot chamber 106 and spring chamber 104 of the load sensing valve 101 are provided with the flow rate adjusting section 70 of the main valve 2.
The front and rear pressures are transmitted to each other. now,
Assuming that the differential pressure before and after the flow rate adjustment section 70 is less than the spring pressure of the spring 105 in the spring chamber lO4 of the load sensing valve 101, that is, a constant pressure (6ky/d), the load sensing valve 101 is located at the symbol position ■2. Then, the discharge amount control section 115 of the variable pump 6 is connected to the pilot line 114,
communicate with the tank 118 via ports n, n+o of the load sensing valve 101 and the pilot line 112;
A swash plate (not shown) of the variable pump 6 is tilted toward the maximum discharge side to increase the discharge amount and increase the differential pressure before and after the flow rate adjustment section 700.

On the other hand, when the pressure in front of the flow rate adjustment section 70 rises and the differential pressure before and after it exceeds the constant pressure, the pressure difference between the pilot chamber 106 of the load sensing 101 and the spring chamber 104 increases When the spring force of the spring 105 of the chamber 104 is exceeded, the load sensing valve 101 moves to symbol position ■1.
Located in the direction.

The discharge amount control section 115 of the variable pump 6 is connected to the pilot line 114, and the port n of the load sensing valve 101 is connected to the pilot line 114.
et, pump line 5 via pilot line 111
8, the swash plate of the variable pump 6 is tilted in the direction of decreasing the discharge amount, thereby reducing the discharge amount and reducing the differential pressure before and after the flow rate adjustment section 70.

To summarize, the load sensing valve 101
Depending on the differential pressure around 00, the variable pump 6 is positioned at the symbol position v or at the symbol position ■2.
The differential pressure before and after the flow rate adjustment section 700 is controlled to be constant by controlling the discharge amount. In other words, in order to avoid discharging unnecessary fluid, the discharge amount of the variable pump 6 is controlled by the load sensing valve 101 so that the differential pressure before and after the flow rate adjustment section 700 is constant. Therefore, this fluid control device is energy saving. Note that in the above operation, the pilot relief valve 102 remains closed.

On the other hand, when the pilot relief valve 102 operates and the pressure in the chamber 104 of the load sensing valve 101 is controlled to the set pressure of the pilot relief valve, the load sensing valve 101 controls the flow rate by the action of the throttle 118. Regardless of the differential pressure before and after the adjustment section 700, the pump line 58
The discharge amount of the variable pump is controlled by positioning it at symbol position ■l or ■2 so that the pressure of and perform pressure control.

The fluid control device configured as described above operates as follows.

Now, the above-mentioned electromagnetic proportional type throttle switching valve 1 is: Lunoid 41a
It is assumed that the hydraulic cylinder 4 is in the state shown in FIG. 1 in which a predetermined amount of current is applied to the hydraulic cylinder 4 and is moving the load 5 to the left in the figure.

At this time, in the spool 23 of the main valve of the throttle switching valve,
Axial thrust W due to fluid flowing through the supply side passage from pump port) P to load port A, and axial thrust W due to fluid flowing through the return side fluid passage from load port) B to tank port T.
2 is given.

The axial thrusts W, W2 are as follows according to the law of conservation of momentum. That is, axial thrust w, -ρQ, V, cosψ1...
...(1) Axial thrust W2-ρQ2 V2 ”'ψ2 ・
・・・・・・・・・(2).

Here, ρ: raw fluid density, Qo: flow rate on the supply side.

Q2: Flow rate on the return side, ■1: Velocity of the fluid in the flow rate adjustment section 70, V2: Velocity of the fluid flowing through the through hole 82.

ψ1: An angle of deviation between the velocity vector of the fluid flowing through the flow rate adjustment unit 70 and the axis of the spool 23, ψ2: An angle of deviation between the velocity vector of the fluid flowing through the through hole 32 and the axis of the spool.

On the other hand, the deflection angle ψ2 is approximately 90° because the through hole 32 is perpendicular to the axis of the spool 23 and is V-shaped. Therefore, the axial thrust W2=ρQ2v2CO892ki0, and the spool 23 stably stands still, causing no jerking in the hydraulic cylinder 4. In other words, due to the jerking phenomenon that would occur if the through hole 32 of this means 30a is not provided, that is, an overrun of the load 5 having large inertia, the return side fluid from the head side port (H) of the hydraulic cylinder 4 may Due to the non-zero axial thrust W2 caused by the return side fluid that rapidly increases and is added to the axial thrust W2 caused by the supply side fluid, the spool 23 moves in the direction of closing all ports and suddenly stops the load 5. On the other hand, when the load suddenly stops, the flow rate becomes zero and the axial thrusts W1, W2
becomes zero, so each port opens again and the load 5 runs again, and this jerking phenomenon occurs due to the resonance between the load sensing valve 101, the variable pump 6, and the throttle switching valve 1. This is not caused by means 30a.

Also, the center of the annular groove 15 of the main valve 2 and the flow rate adjustment part 70 are connected to each other.
If the distance between the annular groove 13 and the through hole 32 is 1, and the distance between the center of the annular groove 13 and the through hole 32 is L2, then the through hole 32 is inevitably closer to the left side than the right end surface 4o of the spool 230 land 2o in the figure, and the distance L2 is becomes large, so that the element ρL2-!, which makes the spool 23 stable in an unsteady state, is relative to the well-known element ρL1 that makes the spool 28 unstable in an unsteady state. ! 3L is
It is larger than i in the past, and this also allows the spool 23 of the main valve 2 to operate stably.

Next, in the throttle switching valve l in the state shown in FIG.
The current value applied to the solenoid 413 is further increased to displace the spool 23 to the left, and the flow rate adjustment section 7
In addition to increasing the opening degree of 0, the land 2 of the spool 28
The right end surface 4o of the hydraulic cylinder 4 is separated from the inner wall 11C of the cylinder chamber 11 and positioned within the annular groove 12, and the hydraulic cylinder 4 is accelerated. Then, when the displacement of the spool 23 from the neutral position is small as shown in FIG. A part of the return side fluid passes through the through hole 82, and another part passes between the right end surface 40 of the land 20 and the inner part 1i111C of the cylinder chamber 11.

Therefore, the pump pressure of the pump 55 of this embodiment with respect to the displacement of the spool of the main valve 2 from the neutral position is
As shown by the curve (A) in the figure, when the displacement of the spool 23 is small, when the return side fluid passes through the through hole 32, the net pressure shown by the curve (B) required to drive the hydraulic cylinder 4 is lower than the net pressure required to drive the hydraulic cylinder 4. When the resistance increases at medium speeds, that is, the back pressure increases at medium speeds as shown by diagonal lines in FIG. 2, and when the displacement of the spool 23 from the neutral position is large, Since the opening degree of is increased, it becomes approximately equal to the above net pressure (b).

This back pressure when the displacement of the spool 28 is small, that is, when the hydraulic cylinder 4 is at medium or low speed, together with the fact that the means 30a does not generate an axial thrust by the return side fluid, prevents the hydraulic cylinder 4 from jerking. This back pressure is generally large at medium and low speeds where the jerking phenomenon is likely to occur, so it is advantageous in preventing the occurrence of jerking, and it is small when the hydraulic cylinder 4 is running at high speeds, so it is preferable from the viewpoint of energy saving.

Note that the curve (C) in FIG. 2 shows the pump pressure of a conventional fluid system in which the return side fluid is returned directly to the tank through a line provided with a fixed throttle without passing through the throttle switching valve. In this case, since the throttle does not function as a variable throttle as in the above embodiment, the pump pressure increases extremely when the hydraulic cylinder is running at high speed, which is undesirable from the viewpoint of energy saving.

In the above embodiment, the means 3a for reducing the axial thrust is provided in the return side fluid passageway leading from the load port B leading to the tank port T, which leads to the head side port (H) of the hydraulic cylinder 4, which has a large flow rate on the return side. However, if necessary, it is of course possible to further provide means for reducing the axial thrust in the return side fluid passage from the load port A to the tank port T.

of Figures 8.4 and 5.6 show △ which reduces the axial thrust, respectively. A modification of the means is shown.

The means 30b shown in FIG. 3 is provided in the means 30a shown in FIG. 1 with notches 85 extending in the axial direction from the end surface 40 of the land 20 at equal intervals on the outer periphery of the land 2.0 to improve variable aperture characteristics. This is what I did.

The means 30C shown in FIG. 4 is such that a sleeve 91 fitted to the main body 1O is provided with radial through holes 92 at equal intervals on the circumference to reduce the axial thrust.

The means 80d shown in FIG. 5 allows the fluid to flow along the curved surface of the groove 96 of the four-land spool 95.
The axial thrust is reduced by making the inclination angle of the axially directed flow of the fluid with respect to the axial center of the spool 95 substantially equal to the inclination angle of the flow directed away from the fluid axis with respect to the spool 95 axis. This is how it was done.

The means 80e shown in FIG. 6 is a 4-land type spool 97 provided with a through hole 32 similar to the means 30a shown in FIG. This reduces the axial thrust.

Moreover, FIGS. 7.8 and 9.10 show modified examples of the power matching circuit, respectively.

The power matching circuit shown in FIG. 7 is constructed by adding a feed-in line 18o having a feed-in throttle 181 to the power matching circuit 7 shown in FIG.
01 with faster response. The power matching circuit shown in FIG. 8 is composed of a load sensing valve 101 and a pressure control valve 150 that controls the discharge amount of the variable pump and regulates the maximum pressure of the pump line. The circuit shown in FIG. 9 is a feed-in aperture 13 to the circuit shown in FIG.
1. A feed-in line 180 having a feed-in line of 1 is provided. The circuit shown in FIG. 10 is the pressure control valve 15 shown in FIG.
The fluid in the pump line is led to the spring chamber of 0 through a throttle 161, and a pilot relief valve 160 is connected to the spring chamber so that the maximum pressure of the circuit can be variably set. Although not shown, the power matching circuit may be of a surplus flow type.

In the above embodiment, a closed center type throttle switching valve has been described, but the present invention is also applicable to an open center type, A,
Needless to say, the present invention can be applied to B and T connection type throttle switching valves, as well as direct-acting type throttle switching valves. Further, the variable pump may have a so-called reverse characteristic in which the discharge amount increases when the pressure transmitted to the discharge amount control section is increased.

As is clear from the above description, according to the present invention, means for reducing the axial thrust is provided in the return side fluid passage of the throttle switching valve having at least four ports, thereby stabilizing the operation of the spool. Therefore, it is possible to prevent the occurrence of the jerking phenomenon caused by the interaction between the variable pump, load sensing valve, throttle switching valve, and actuator that drives a load with large inertia.

In addition, since the discharge amount of the variable pump is controlled according to demand by a load sensing valve, there is no wasted discharge fluid, and the means for reducing the above-mentioned axial thrust is not a throttle as in the conventional example, so This has the advantage that the back pressure of the fluid on the return side is reduced, and therefore the energy loss is reduced. Further, since the means for reducing the axial thrust is provided in the return side fluid passage of the throttle switching valve having at least four ports, the direction control and speed control of the actuator can be easily and accurately controlled by the throttle switching valve. Therefore, there is an advantage that the circuit configuration is simplified.

[Brief explanation of the drawing]

Fig. 1 is a circuit diagram showing a part of a fluid control device according to an embodiment of the present invention in cross section, Fig. 2 is a graph showing the relationship between the displacement of the spool of the main valve and the pump pressure, and Fig. 8.4. 5.6
The figures are a cross-sectional view showing means for reducing the axial thrust, and a seventh figure.
．． Figures 8, 9, and 10 are circuit diagrams of modified examples of the power matching circuit, respectively. 1... Throttle switching valve, 2... Main valve,
8...Pilot valve, 4...Actuator, 5...Load, 6...
Variable pump, 28.95.97...Spool, 80a, 80b, 80c, Sod. 80e... Means for reducing axial thrust, 101
...Load sense sink valve. Patent Applicant: Daikin Industries, Ltd. 2-114

Claims (1)

[Claims] (1) Variable bonder (6) and actuator (4)
, a throttle switching valve (1) having at least four boats connected between the variable pump (6) and the actuator (4), and a discharge amount control section (11) of the variable pump (6).
5) A load sensing valve (
101), and controls the discharge amount of the variable pump (6) by making the load sensing valve (ioB respond to the differential pressure before and after the flow rate adjustment part of the throttle switching valve (1)) to switch the throttle. In a fluid control device that controls the differential pressure before and after the flow rate adjustment part of the valve (1) to a constant value, and controls the direction and speed of the actuator (4) by the throttle switching valve (1), the throttle switching valve (1) controls the direction and speed of the actuator (4). Means (80) for reducing the axial thrust of the fluid on the spool (28) K is provided in at least one return side fluid passage of the valve (1), and the operation of the spool (28) is stabilized, and the actuator (4) ) is a fluid control device that is characterized by being able to drive smoothly without jerking even with loads with large inertia.