JPH0319401B2 - - Google Patents

Description

Detailed Description of the Invention (Field of Industrial Application) The present invention is suitable for use in, for example, a mold opening/closing device of an injection molding machine, and can speed up an actuator that drives a load with large inertia smoothly without so-called shakiness. The present invention relates to a fluid control device that can be driven while being controlled.

(Prior Art) Generally, in a meter-in type fluid control device that adjusts the operating speed by restricting the flow rate in the inlet side conduit of the actuator, it is known that sagging occurs in the actuator.

For example, when an actuator that drives a load with large inertia, such as a mold in an injection molding machine, is controlled to decelerate from high speed to low speed, the actuator will overrun (run by itself at its initial speed) due to the inertia of the mold. At this time, the pressure in the inlet side conduit of the actuator suddenly drops below the driving pressure of the actuator, and when the actuator stops, the pressure in the inlet side conduit rises again to drive the actuator. As a result of repeating such an operation, the actuator becomes eroded.

When this sagging occurs, the equipment itself is damaged and its commercial value is lost.As a solution to this problem, a device as shown in Figure 8 has been proposed for injection molding machines (Hydraulic and Pneumatic Design, Vol. 15, Vol. No. 10, no.
39 pages). This device consists of two pumps 155, 15
6, and a fixed throttle 1 is connected to the tank line 153 of the directional valve 151.
54 and a pilot check valve 159 which is opened and closed by a switching valve 158 are connected in parallel.
9 is opened to prevent back pressure from being generated in the tank line 153. When deceleration control is performed from high speed to low speed, by closing the pilot check valve 159, back pressure is applied by the fixed throttle 154 to the return line of the actuator, so overrun of the actuator 152 is prevented, and deceleration control without shaking is performed. It is possible.

(Problems to be Solved by the Invention) However, since this device uses various devices such as a fixed throttle 154, a pilot check valve 159, and a switching valve 158 for opening and closing this check valve, the circuit becomes complicated. The deceleration operation becomes complicated. Further, the pressure of the pump line is always the set pressure of the relief valve 157, and even at this pressure, surplus flow is released to the tank, so there is a drawback that power loss is large.

An object of the present invention is to eliminate the above-mentioned drawbacks by using a variable pump to reduce energy loss while suppressing its vibration and noise, and to provide an actuator having an inertial load with a simple circuit configuration. An object of the present invention is to provide a fluid control device capable of smoothly operating a fluid controller without any lag.

(Means for solving the problem) For this reason, the fluid control device of the present invention has a spool 23 including a first land 21 and a second land 20, and the spool 23 can be displaced in the axial direction. Therefore, a first passage 70 connecting between the first load port A and the pump port P and a second passage 30a connecting between the second load port B and the tank port T are formed by using the end surfaces of the lands 21 and 20. At the same time, after opening, the spool 2
Both passages 70, 30 are set to values corresponding to the displacement amount of 3.
A pilot valve type 4-port throttle switching valve 1 whose opening degree can be adjusted, and a discharge amount control section 115 of the variable pump 6 in response to the differential pressure before and after the first passage 70.
and a load sensing valve 101 which controls the discharge amount of the variable pump 6 by switching communication with the pump line 53 and the tank 113, and the spool 2 in the 4-port throttle switching valve 1.
The second land 20 of No. 3 is a throttle passage 31 which connects the end face of the passage opening/closing side of the second land 20 and the outer peripheral surface of the second land 20, and whose opening direction to the outer peripheral surface is the radial direction. 32, and when the opening degree of the first passage 70 is reduced by the displacement of the spool 23 to achieve a low flow rate, the end face of the second land 20 allows the gap between the second load port B and the tank port T to be reduced. The second passage 30a is closed, and the second load port and tank port T are communicated only by the throttle passages 31 and 32, so that back pressure is applied.

(Function) By displacing the spool 23 of the throttle switching valve 1 using the pilot valve, the connection between the first passage 70, which communicates between the first load port A and the pump port P, and the second load port B, and the tank port T is established. The communicating second passage 30a is opened, and the degree of opening of both passages 70, 30a is adjusted.

At this time, pressure is compensated to a constant pressure difference before and after the first passage 70 by the load sensing valve 101.
The load sensing valve 101 regulates the discharge amount of the variable pump 6 to a value corresponding to the opening degree of the first passage 70. Therefore, a flow rate proportional to the opening degree of the first passage 70 is supplied to the actuator. On the other hand, the second passage 30a and the throttle passages 31 and 32 between the second load port B and the tank port T are open, and the degree of opening of the second passage 30a is larger than that of the throttle passages 31 and 32. Part of the fluid flows through the second passage 3
Since the flow rate passing through the throttle passages 31 and 32 through 0a is small, the discharge flow rate released from the actuator is released into the tank with only slight throttle resistance (negligible). It is assumed that the opening degree of the first passage 70 is reduced by the spool 23 and the second passage 30a is closed. At this time, the actuator attempts to overrun (self-propelled) due to the inertia of the load. However, since the discharge flow rate from the actuator is subject to restriction resistance by the restriction passages 31 and 32, back pressure is generated in the return line of the actuator, and this back pressure applies a braking force to the actuator, thereby preventing overrun of the actuator. During this stable deceleration without any deceleration, the pressure downstream of the first passage 70 does not drop. The pressure increases by an amount corresponding to the braking force, and the degree of increase is alleviated by adjusting the opening degree from the second passage to the throttle passage. Therefore, the load sensing valve 101 controls the variable pump 6 so that the pressure in front of the first passage 70 is increased by the pressure corresponding to the braking force in a stable state, and after deceleration, the pressure in the front part of the first passage 70 is increased by the pressure corresponding to the braking force.
The actuator is driven with a flow rate proportional to the opening degree of the passage 70.

(Embodiments) The present invention will be described in detail below with reference to illustrated embodiments.

In Fig. 1, 1 indicates the main valve 2 and the pilot valve 3.
4 is a hydraulic cylinder that drives a load 5 with large inertia, 6 is a variable pump, and 7 is a power matching circuit that controls the discharge amount of the variable pump 6.

The main valve 2 is a closed center type 3-position switching valve, and has a cylinder chamber 11 formed in the main body 10, and annular grooves 12, 1 are formed in the cylinder chamber 11 sequentially from the left side.
3, 14, 15, and 16 are provided. The annular groove 12 communicates with the tank port T, the annular groove 13 communicates with the load port B, the annular groove 14 communicates with the pump port P, the annular groove 15 communicates with the load port A, and the annular groove 16 communicates with the tank port T. . The inner wall 11b of the cylinder chamber 11 between the annular grooves 13 and 14 has an opening 18b of the feedback passage 17, and the inner wall 11a between the annular grooves 14 and 15 has an opening 18a of the feedback passage 17. has been established.

The cylinder chamber 11 has three lands 20, 2.
A spool 23 having numbers 1 and 22 is slidably fitted, and the pump port P is communicated with the load port A or B and the tank port T is communicated with the load port B or A depending on the position of the spool 23. .

Then, depending on the displacement of the spool 23, the opening degree of the first passage 70 that communicates between the pump port P and the first load port A, and the opening degree of the second passage 30a that communicates between the second load port B and the tank port T. Adjust. On the other hand, the land 20 of the spool 23 is provided with an axial annular groove 31 that opens on the right end surface 40 of the land 20 and a radial through hole 32 that opens on the outer peripheral surface of the land 20. 32 and the annular groove 31 are connected to form a throttle passage. The throttle passages 31 and 32 communicate between the load port B and the tank port T when the second passage 30a that communicates between the load port B and the tank port T is closed by the end face of the land 20.

Furthermore, pilot chambers 25 and 26 are formed at both ends of the spool 23, respectively.
Bias coil springs 2 at 5 and 26 respectively
7 and 28 are inserted.

On the other hand, the pilot valve 3 is an electromagnetic pressure reducing valve having a well-known switching function, and has solenoids 41a and 41b.
, a primary port m, and secondary ports _nA , _nB , and controls the fluid pressure of the secondary port nA or nB to be reduced to a pressure proportional to the current value applied to the solenoid 41a or 41b, The secondary port _nB or _nA on the side not subjected to pressure reduction control is communicated with 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 _nB is connected to the pilot chamber 25 of the main valve 2 via a line 45. Also, secondary port _nA is connected to pilot chamber 26 via line 46.

Therefore, if current is applied to the solenoid 41a or 41b of the pilot valve 3 in the pilot chamber 26 or 25 of the main valve 2, its secondary port
The pressure of n _A or n _B is transmitted to the spool 2 above.
3 is displaced to a position where the pressure transmitted to one end surface and the spring force of the other coil spring 27 or 28 are balanced, and the opening degrees of the first passage 70 and the second passage 30a are proportional to the above current value. It has come to be determined that

On the other hand, the load port A of the main valve 2 is connected to the line 51.
The load port B is connected to the rod side port R of the hydraulic cylinder 4 via a line 52, and the load port B is connected to the head side port H of the hydraulic cylinder 4 via a line 52. 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 and T of main valve 2 are connected to tank 5, respectively.
6,56.

On the other hand, the power matching circuit 7 includes a 3-port pilot valve 10 as an example of a load sensing valve.
1 and a pilot relief valve 102.

The above load sensing valve 101 is at the symbol position
Ports l and n are communicated at V ₁ and port m is closed, while port m and port n are communicated at symbol position V ₂ and port l is closed. For example, the spring force of the spring 105 is set to correspond to a differential pressure of 6 kg/cm ² , and when the differential pressure between the pilot chamber 106 and the spring chamber 104 is 6 kg/cm ² or more, the symbol position V ₁ is set.
, and when the differential pressure is 6 kg/cm ² or less, it is positioned at symbol position V ₂ .

A pump line 53 is connected to port l of the load sensing valve 101 via a pilot line 111, and a tank 113 is connected to port m via a pilot line 112. A variable pump 6 is connected to port n of the load sensing valve 101 via a pilot line 114.
A discharge amount control section 115 consisting of a swash plate control cylinder that is always biased toward the maximum discharge side by a spring, for example.
Connect. Furthermore, the load sensing valve 10
Pilot line 11 in the pilot compartment 106 of 1
The pump line 53 is connected to the spring chamber 104 via the valve 6, while the spring chamber 104 is connected to the field back passage 17 of the main valve 2 via a pilot line 119 having a throttle 118. Between the spring chamber 104 and the throttle 118 of the load sensing valve 101,
A tank 121 is connected via a pilot line 120 provided with an intermediate pilot relief valve 102.

Therefore, the pressure before and after the first passage 70 of the main valve 2 is transmitted to the pilot chamber 106 and the spring chamber 104 of the load sensing valve 101, respectively. Now, the differential pressure before and after the first passage 70 is the spring pressure of the spring 105 in the spring chamber 104 of the load sensing valve 101, that is, a constant pressure (6 kg/cm ² ).
If the load sensing valve 101 is less than
is located at the symbol position _V2 , and the discharge amount control section 115 of the variable pump 6 is connected to the pilot line 114,
The first passage 70 is connected to the tank 113 through ports n and m of the load sensing valve 101 and the pilot line 112, and the swash plate (not shown) of the variable pump 6 is inclined toward the maximum discharge side to increase the discharge amount. Increase the differential pressure before and after.

On the other hand, when the pressure at the front side of the first passage 70 rises and the differential pressure before and after it becomes below the constant pressure, the pilot chamber 1 of the load sensing valve 101
06 and the spring chamber 104, the pressure difference between the spring chamber 1
04, the load sensing valve 101 is positioned in the symbol position _V1 direction. Then, the discharge amount control section 11 of the variable pump 6
5 to the pilot line 114, ports n and l of the load sensing valve 101, and the pilot line 11
1 to the pump line 53, the swash plate of the variable pump 6 is tilted in the direction of decreasing the discharge amount, thereby reducing the discharge amount and the differential pressure across the first passage 70. To summarize, the load sensing valve 101 is positioned at the symbol position V ₁ or at the symbol position V ₂ depending on the pressure difference before and after the first passage 70 to control the discharge amount of the variable pump 6. Well, the first
The differential pressure across the passage 70 is controlled to be constant. In other words, in order to avoid discharging wasteful fluid, the discharge amount of the variable pump 6 is controlled by the load sensing valve 101 so that the differential pressure across the first passage 70 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, if the pilot relief valve 102 is operated to control the pressure in the spring chamber 104 of the load sensing valve 101 to the set pressure of the pilot relief valve, the load sensing valve 101 Regardless of the differential pressure before and after the valve 70, the pressure in the pump line 53 is lower than the set pressure of the pilot relief valve 102 in the spring chamber 1.
Pressure control is performed by controlling the discharge amount of the variable pump 6 by positioning it at the symbol position V ₁ or V ₂ so as to increase the spring pressure of the spring 105 of 04.

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

Now, in the electromagnetic proportional throttle switching valve 1, a predetermined amount of current is applied to the solenoid 41a, and the hydraulic cylinder 4 is in the state shown in FIG. 1 moving the load 5 to the left in the figure. do.

At this time, the spool 2 of the main valve of the throttle switching valve
3 is given an axial thrust W ₁ by the fluid flowing in the supply side passage from the pump port P to the load port A, and an axial thrust W ₂ due to the fluid flowing in the return side fluid passage from the load port B to the tank port T. It will be done.

The axial thrusts W ₁ and W ₂ are as follows according to the law of conservation of momentum. That is, axial thrust W ₁ = ρQ ₁ V ₁ cosψ ₁ ………(1) Axial thrust W ₂ = ρQ ₂ V ₁ cosψ ₂ ………(2).

Here, ρ: density of liquid, Q ₁ : flow rate on supply side,
Q ₂ : Flow rate on the return side, V ₁ : Velocity of the fluid in the first passage 70, V ₂ : Velocity of the fluid flowing in the through hole 32,
ψ ₁ : Deviation angle between the velocity vector of the fluid flowing through the first passage 70 and the axis of the spool 23, ψ ₂ : Through hole 3
This is the angle of deviation between the velocity vector of the fluid flowing through 2 and the axis of the spool.

On the other hand, since the through hole 32 is perpendicular to the axis of the spool 23, the deviation angle ψ ₂ is as follows.
It will be approximately 90°. Therefore, axial thrust W ₂ =ρQ ₂ ,
Since V ₂ cosψ ₂ ≒0, the spool 23 is stably stationary and the hydraulic cylinder 4 does not sag. In other words, due to the sagging phenomenon that occurs when the through hole 32 is not provided, that is, due to an overrun of the load 5 having a large inertia, the return side fluid from the head side port H of the hydraulic cylinder 4 increases rapidly. Due to the non-zero axial thrust W ₂ caused by the return side fluid added to the axial thrust W ₁ caused by the supply side fluid, the spool 23 moves in the direction of closing all ports and suddenly stops the load 5. When it stops, the flow rate becomes zero and the axial thrusts W ₁ and W ₂ become zero, so each port opens again and the load 5 overruns again, and the load sensing valve 101, variable pump 6 and throttle The through hole 32 prevents the sagging phenomenon caused by resonance with the switching valve 1 from occurring.

Further, if the distance between the center of the annular groove 15 of the main valve 2 and the first passage 70 is L ₁ and the distance between the center of the annular groove 13 and the through hole 32 is L ₂ , then the through hole 32 is connected to the spool 23 . Since the land 20 inevitably moves to the left side of the right end surface 40 in the figure, the distance L ₂ becomes larger.
The element ρL ₂ ×dQ ₁ /dt that stabilizes the spool 23 in an unsteady state is larger than that in the past, compared to the well-known element ρL ₁ ×dQ ₁ /dt that makes the spool 23 unstable in an unsteady state. Even if it sways, the spool 23 of the main valve 2 is stably operated.

Next, in the throttle switching valve 1 in the state shown in FIG. 1, the current value applied to the solenoid 41a is further increased, the spool 23 is displaced to the left, and the opening degree of the first passage 70 is increased. At the same time, the right end surface 40 of the land 20 of the spool 23 is separated from the inner wall 11c of the cylinder chamber 11, and the annular groove 1
2 and opens the second passage 30a to increase the speed of the hydraulic cylinder 4. Then, as shown in FIG. 1, when the displacement of the spool 23 from the neutral position is small, the return side fluid passes only through the through hole 32, whereas as described above
When the displacement of
0 and the inner wall 11c of the cylinder chamber 11 through a second passage 30a.

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 curve A in FIG. The net pressure required for driving, shown by curve B, is greater than the resistance required at medium speeds when the return fluid passes through the through hole 32, that is, at medium speeds, as shown by diagonal lines in FIG. When the back pressure is increased by the increased back pressure and the displacement of the spool 23 from the neutral position is large, the opening degree of the return side fluid passage is increased, so that the net pressure becomes approximately equal to the net pressure (b).

This back pressure when the displacement of the spool 23 is small, that is, at medium and low speeds of the hydraulic cylinder 4, together with the fact that the through hole 32 does not generate an axial thrust by the return side fluid, prevents the hydraulic cylinder 4 from sagging. This back pressure is generally large at medium and low speeds where the sagging phenomenon is likely to occur, so it is preferable in order to prevent the occurrence of sagging.
Since it is small at high speeds, it is also preferable from the viewpoint of energy saving.

Note that 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 throttle passages 31 and 32 are provided in the return passage leading from the load port B to the tank port T, but if necessary, the throttle passage 31 and 32 are provided in the return passage leading from the load port A to the tank port T.
32 may of course be provided.

As shown in FIG. 3, notches 85 extending in the axial direction from the end surface 40 of the second land 20 are provided at equal intervals on the outer periphery of the land 20 to improve the variable aperture characteristic.

Further, FIGS. 4, 5, 6, and 7 show modified examples of the power matching circuit, respectively.

The power matching circuit shown in FIG. 4 is a power matching circuit 7 shown in FIG.
A feed-in line 130 having a diameter of 31 is provided to speed up the response of the load sensing valve 101. The power matching circuit shown in FIG. 5 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. 6 is obtained by adding a feed-in line 130 having a feed-in aperture 131 to the circuit shown in FIG. The circuit shown in FIG. 7 leads the fluid of the pump line to the spring chamber of the pressure control valve 150 shown in FIG. 6 through a throttle 161, and connects a pilot relief valve 160 to variably set the maximum pressure of the circuit. It has been made possible.

In the above embodiment, a closed center type throttle switching valve has been described, but the present invention is applicable not only to open center type and A, B, T connection type throttle switching valves, but also to 2-position throttle switching valves. It goes without saying that it can be done.

(Effects) As is clear from the above description, according to the present invention, by only operating the spool 23, the second passage 30a between the second load port B and the tank port T,
Since back pressure is applied or not applied to the return line of the actuator by selecting the throttle passages 31 and 32, the circuit configuration and operation are simpler than in the past. Furthermore, when decelerating the actuator, back pressure is generated in the return line in the radial throttle passage 32, which prevents axial thrust from being generated on the spool and prevents overline due to inertia of the actuator. Stable deceleration can be obtained.

Moreover, since the opening degree of the second passage 30a can be adjusted by the spool, the switching shock when decelerating is eased.

Furthermore, the operation of the variable pump and the load sensing valve is stabilized due to the function of the throttle passage, and the vibration and noise of the variable pump are similarly suppressed. Since only the flow rate corresponding to the amount of water is discharged and no excess flow rate is discharged, there is no power loss and it has the advantage of being excellent in energy saving effect.

[Brief explanation of drawings]

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 displacement of the spool of the main valve and pump pressure, and FIG. A sectional view showing a modification of this invention,
4, 5, 6, and 7 are circuit diagrams of modified examples of the power matching circuit, respectively, and FIG. 8 is a diagram showing a conventional flow rate control device. 1... Throttle switching valve, 2... Main valve, 3... Pilot valve, 4... Actuator, 5... Load, 6
...Variable pump, 23...Spool, 30a, 3
0b...Second passage, 31, 32...Aperture passage, 7
0...First passage, 101...Load sensing valve.

Claims (1)

[Claims] 1. A spool 23 including a first land 21 and a second land 20 is provided, and by displacing the spool 23 in the axial direction, both lands 21,
A first passage 70 which communicates between the first load port A and the pump port P, and a second passage 3 which communicates between the second load port B and the tank port T, has an end face of 20.
0a, and after opening, the pilot valve type 4-port throttle switching valve 1 is capable of adjusting the opening degree of both the passages 70 and 30a to a value corresponding to the displacement amount of the spool 23; 1 aisle 7
In response to the differential pressure around 0, the discharge amount control unit 115 of the variable pump 6 is controlled between the pump line 53 and the tank 11.
and a load sensing valve 101 that controls the discharge amount of the variable pump 6 by switching communication with the spool 2 in the 4-port throttle switching valve 1.
The second land 20 of No. 3 is a throttle passage 31 which connects the end face of the passage opening/closing side of the second land 20 and the outer peripheral surface of the second land 20, and whose opening direction to the outer peripheral surface is the radial direction. 32, and when the opening degree of the first passage 70 is reduced by the displacement of the spool 23 to achieve a low flow rate, the end face of the second land 20 allows the gap between the second load port B and the tank port T to be reduced. A fluid control device characterized in that the second passage 30a is closed and the second load port B and the tank port T are communicated only by the throttle passages 31 and 32 to apply back pressure.