JPH04302784A - Flow control valve - Google Patents

Abstract

PURPOSE:To enable easy and accurate flow control of a refrigerant through controlling the value of an electric current applied to an electromagnetic coil, by applying an electric current to the electromagnetic coil, thereby attracting a plunger, and thereby opening a valve member against the forces of springs. CONSTITUTION:Applying an electric current to an electromagnetic coil 1 causes the formation of a magnetic circuit composed of a plunger 13, an attractor 10, a housing 18 and a cap 20, the resulting magnetic flux attracting the plunger 13 to the attractor 10. Then, a rod 11 moves a valve member 3 in a valve- opening direction against the forces of springs 6. At this time, the valve member 3 is moved up to a position where the moved amount thereof is balanced with a load obtained by multiplying a composite spring constant K of the springs 16 and 15 by the mentioned moved amount L of the valve member 3. As a result, a valve seat 27 is opened to permit a high-pressure refrigerant to flow into a low-pressure refrigerant passage 5 via a valve orifice 6. Namely, since the position for a balance between the attractive force and the load obtained through multiplication of the values KXL continuously changes, the area of the opening formed by the valve seat 27 and the valve member 3 continuously changes.

Description

[Detailed description of the invention]

0001

[Industrial Application Field] The present invention relates to a flow control valve used in the refrigeration cycle of car air conditioners, vehicle-mounted air conditioners, room air conditioners, package air conditioners, and these heat pump air conditioners. The present invention relates to a flow rate control valve that can arbitrarily set the flow rate of refrigerant and is suitable for performing a wide range of control regardless of the type of refrigerant.

0002

2. Description of the Related Art Conventionally, various flow control valves have been proposed that utilize electromagnetic valve technology to continuously control the flow rate of refrigerant in response to an electrical input signal. Various conventional examples will be cited and explained below.

1. The first conventional example is shown in FIG.
2-45651 ``Expansion valve''.

[0004] This expansion valve is of a duty drive type in which the electric input signal applied to the electromagnetic coil 1 is a rectangular current, and the energization ratio within one cycle is varied.

In this system, a valve body 3 slidably disposed on an expansion valve body 2 moves up and down in response to an input signal, opening and closing a valve orifice 6 disposed on a side wall per unit time. As a result, the flow rate of the refrigerant is continuously controlled from the high-pressure refrigerant flow path 4 to the low-pressure refrigerant flow path 5, and there is no flow hysteresis, and the structure is simple, compact, lightweight, and inexpensive.

2. A second conventional example is shown in FIG.
-38778 "Electromagnetic expansion valve in cooling circuit" is cited and described.

This expansion valve is of a current value control type in which the electrical input signal applied to the electromagnetic coil 1 is a direct current. In this method, a valve body 3, which is integrated with a plunger 13 that is in contact with a valve orifice 6 by a spring 15, is slidably disposed on an expansion valve body 2, and this valve body is caused by direct current. Due to the magnetic force, it is attracted to the attractor 10 side to a position where it balances the pressing force of the spring 15,
Open the valve orifice 6. In other words, this is a method of controlling the position (stroke control) of the valve body 3 relative to the valve orifice 6 by moving the valve body 3 by changing the current value.

3. A third conventional example is ``Reversible Expansion Valve'' shown in JP-A-62-194081 shown in FIG.

In the third conventional example, the high pressure refrigerant flow path 4
There are two holes orthogonal to each other, and the upper part in the figure is the sliding hole 9 for the differential pressure elimination rod 25, and the lower part is the valve orifice 6. Further, a pressure equalizing hole 8 is provided that communicates the low pressure refrigerant flow path 5 with the plunger chamber 26 . Therefore, the balance of forces is: cross-sectional area of the differential pressure elimination rod 25: A1, cross-sectional area of the valve orifice A0, high pressure: PH, low pressure: PL, electromagnetic force: F, pressing force of the spring 15: K1, spring 16. Pressing force: K2, then F+(PH x A1)+(PL x A0)-(PL
×A1 )-(PH ×A0)-K1 +K2 =
It becomes 0.

[0010] Here, since A1 = A0, F
= K1 - K2 Therefore, the position control (stroke control) of the valve body 3 can be performed according to the DC current value without being affected by the pressure fluctuation of the refrigerant, and therefore the refrigerant flow rate accuracy is good.

[0011]

[Problems to be Solved by the Invention] In the first conventional example, since the valve body repeatedly opens and closes the flow port at a high speed of several tens of Hz,
The disadvantages are that the liquid refrigerant causes a water hammer phenomenon, which causes the piping to resonate and generates abnormal noise, and that the valve stop performance when closing the valve is poor. Further, since the valve body moves up and down at high speed, problems such as wear occur and durability is poor.

In the second conventional example, the valve body has the following formula: (high pressure PH - low pressure PL) x valve orifice cross-sectional area A
A force of 0 = FB (kgf) is always applied upward in the diagram. Since the high pressure during the refrigeration cycle constantly fluctuates, FB also fluctuates, and the position (stroke) of the valve body 3 becomes unstable depending on the current value, making it impossible to accurately control the flow rate. Furthermore, in this method, since flow rate hysteresis exists as is well known, the accuracy of flow rate control is increasingly deteriorated.

[0013] Furthermore, since the high pressure during the refrigeration cycle increases to a maximum of about 30 kgf/cm2 G, A0
According to the calculation when =0.1cm2, FB =30×
0.1=3kgf. Therefore, in order for the valve body to maintain the valve closed state and have the valve sealing function, the spring 15
Since it is necessary to set the pressing force to about 3.5 kgf, a large electromagnetic force is required to open the valve body 3. Therefore, the electromagnetic coil becomes large, making it unsuitable for practical use.

Furthermore, in the third conventional example, because the gap 24 (magnetic path gap) between the plunger 13 and the attractor 10 cannot be made the same between products due to the structure, the electromagnetic force that is greatly affected by this gap is Since the flow rate differs between products, uniform flow rate accuracy cannot be obtained. Furthermore, since this gap becomes the full stroke of the valve, uniform flow rate accuracy cannot be obtained, which is a problem from a quality standpoint.

Furthermore, in order to completely bring the valve body 3 into close contact with the valve orifice 6, it is necessary to set (play) the stroke of the plunger 13 to be greater than the stroke of the valve body 3, and when the maximum current is applied, the plunger 13 and the suction Since the child 10 is configured to be in close contact with each other, it also has the disadvantage of increased flow hysteresis.

When the valve is closed, high-pressure refrigerant passes through the differential pressure elimination rod 25 through the gap in the sliding hole 9 and flows out from the pressure equalizing hole 8 into the low-pressure refrigerant flow path 5.

[0017]

[Means for Solving the Problems] In order to achieve the above object, in the present invention, a through hole including a low pressure side circulation hole orthogonal to a high pressure side circulation hole is provided in the expansion valve body, and one side is connected to the valve body. One of the sliding holes is used as a valve orifice, and a valve body is disposed in this through hole so as to be movable up and down, having a cylindrical shaft shape and a conical shape, and having a pressure equalizing hole penetrated through the shaft center.

Furthermore, the valve body is driven by an electromagnetic mechanism provided in a housing fixed to the valve body, and a cylindrical shaft made of a non-magnetic material is installed in the plunger in order to adjust the usable range of electromagnetic force in the electromagnetic mechanism. Arrange the rod and spring pin,
A hollow, thin disk spacer made of a non-magnetic material was placed between the opposing surfaces of the plunger and the attractor to regulate the amount of movement of the valve body.

The influence of the refrigerant pressure acting on the valve body constructed as described above is as follows: PH is the saturation pressure at the condensation temperature of the refrigerant PL is the saturation pressure at the evaporation temperature of the refrigerant C1 is the flow coefficient θ is the conical part of the valve body. The half apex angle L of the triangular cross section that includes the central axis of the cone with the upper apex angle is the distance that the valve element moves when opening in the axial direction from the valve closing point (valve stroke) D1 is the diameter of the valve orifice D2 is the valve diameter The diameter FO of the cylindrical shaft of the body is given by the force FC acting on the valve body in the valve opening direction and the force acting on the valve body in the valve closing direction.

Since D1 = D2 here, L=0 when the valve is closed, so FO =FC, and the valve body is not affected by the refrigerant pressure at all.

According to this equation, in order to reduce the axial thrust, it is necessary to set the half apex angle θ of the conical portion of the valve body and the valve stroke L small.

FIG. 3 is an example of an experiment in which the degree of influence of the axial thrust acting on the valve body on the valve stroke was conducted using the conical half-apex angle θ of the valve body as a parameter.

According to this, the valve stroke reduction rate reaches its maximum when the cone half apex angle θ is 45°, and reaches 0 when the cone half apex angle θ is 19°. Note that the stroke reduction rate represents the ratio of the valve stroke reduction amount when high pressure is applied to the valve stroke amount when no pressure is applied as 100.

This relationship also changes depending on the clearance S shown in FIG. Since the generation of thrust can be suppressed, the valve body
It will not receive any power.

Therefore, the valve stroke is independent of the refrigerant pressure and is determined only by the electromagnetic force and the pressing force of the spring.

[0026] Therefore, in order to perform accurate position control,
The pressing force of the spring can be adjusted, and a spring constant K is used that prevents fluctuations as much as possible.

However, the magnetic force generated from the electromagnetic coil is
As shown in FIG. 4, the magnetic force varies greatly depending on the size of the gap between the plunger and the attractor, so it is necessary to provide an adjustment structure to make the size of the gap uniform.

FIG. 4 shows the change in magnetic force at each current value using the air gap as a parameter.

FIG. 5 shows the point where the spring displacement load diagram drawn in FIG. 4 intersects the magnetic force curve for the air gap for each current value (this point is the plunger position control point (valve lift point))
is plotted.

As shown in FIG. 4, when the gap size changes as shown in FIG.
P1, P2, P3) Since the current-valve lift characteristics vary greatly as shown in FIG. 5, the accuracy of flow rate control deteriorates.

Therefore, in the present invention, the gap is adjusted by the method shown in FIGS. 6(a) and 6(b). Specifically, as shown in FIG. 6(a), after placing the conical portion of the valve body and the valve seat 27 in close contact with each other, and placing a jig having a thickness of t1 on the end face of the suction element 10, the plunger 13 is inserted. Place it on this jig. Furthermore, the rod 11 is inserted until it contacts the upper end surface of the valve body 3, and the spring pin 14 is press-fitted.

This state determines the P2 point in FIG. Thereafter, as shown in FIG. 6(b), the jig having a thickness of t1 is removed, and a spacer 12 made of a non-magnetic material and having a thickness of t2 is provided, which is thinner than this jig. As a result, point P4 in FIG. 4 is determined. In other words, the difference in thickness between the jig and the spacer (t1 - t2) is set as the valve stroke L.

[0033] According to such a configuration, therefore, there is no variation in the air gap and, moreover, there is no variation in the valve full stroke, making it possible to achieve high flow rate control accuracy.

[0034]

[Operation] The flow control valve of the present invention operates as follows.

The diameter D1 of the valve orifice and the diameter D2 of the cylindrical portion of the valve body are formed to be the same diameter, and the pressure equalization hole is provided in the valve body, so that the pressure in the high-pressure refrigerant flow path and the low pressure are the same. The pressure within the refrigerant flow path acts equally on the valve body in opposite directions in the axial direction. Also, the half apex angle θ of the conical part of the valve body is 3
Since the angle is set to 0° or less, the valve body is not affected by these pressures in a dynamic state as well as in a static state.

Therefore, the valve body is operated by the suction force acting in the valve opening direction, the pressing force of the spring 15, and the spring 1 acting in the valve closing direction.
The position (valve opening degree) is determined by the pressing force of 6. Furthermore, since the suction force and the spring pressing force are each adjustable, highly accurate position control (valve stroke control) is possible.

[0037]

DESCRIPTION OF THE PREFERRED EMBODIMENTS In FIG. 1, reference numeral 2 denotes a main body of a valve, in which a high-pressure refrigerant passage 4, a sliding hole 9 and a valve orifice hole 6 of the same diameter are arranged perpendicularly to the high-pressure refrigerant passage 4. Reference numeral 3 denotes a valve body consisting of a cylindrical shaft-shaped part that is slidably inserted in the axial direction into the sliding hole 9 with a very small gap, and a conical shaft-shaped part that comes into contact with the valve seat 27. A pressure equalizing hole 8 is provided.

7 is a packing made of a low friction material (such as Teflon) that seals the refrigerant flowing out from the extremely small gap formed between the sliding hole 9 and the cylindrical shaft of the valve body, and is fixed by a bush 22. Ru.

A spring 16 for driving the valve body 3 in the valve closing direction is disposed in the low pressure refrigerant passage 5 of the valve body 2.
An adjustment screw 17 is provided to adjust the pressing force of this spring 16.

Reference numeral 13 denotes a cylindrical shaft-shaped plunger made of a magnetic material, which is disposed so as to be slidable in the axial direction in a low pressure chamber formed by an attractor 10 made of a magnetic material, a pipe 19 made of a non-magnetic material, and a spring ke 21. has been done.

Inside the plunger 13, a rod 11 made of a non-magnetic material passes through an internal hole of the attractor 10 and is placed on the cylindrical axis of the valve body 3 in order to provide an appropriate gap on the surface facing the attractor 10. It is fixed by a spring pin 14 at a position where it abuts the end face.

The plunger 13 and rod 11 may be positioned and fixed by press-fitting the rod into the plunger. Alternatively, they may be connected by screws.

Reference numeral 12 denotes a hollow thin disk-shaped spacer made of a non-magnetic material for setting the lift amount of the valve body 3.

Reference numeral 15 denotes a spring that presses the plunger 13, and has the purpose of adjusting the composite spring constant K that is combined with the spring 16.

Reference numeral 18 denotes a hollow cylindrical housing made of a magnetic material, and is mechanically coupled to the valve body 2. The electromagnetic coil 1 is disposed within the housing 18, and is fixed by mechanically coupling a cap 20 made of a magnetic material to the housing 18.

Reference numeral 23 denotes an O-ring that prevents low-pressure refrigerant from leaking to the outside.

In the state shown in FIG. 1, the valve body 3 is pressed against the valve seat 27 by the pressing force of the spring 16, so that the high-pressure refrigerant flows from the high-pressure refrigerant flow path 4 through the valve orifice 6 to the low-pressure refrigerant flow path 5. Refrigerant is shut off.

The operation of the flow control valve constructed as described above will be explained below. Electrical input signal to electromagnetic coil 1 (
When a current) is applied, a magnetic circuit is formed by the plunger 13 - the attractor 10 - the housing 18 - the cap 20, and the magnetic flux (attractive force) attracts the plunger 13 to the attractor 10. Therefore, the rod 11 mechanically connected to the plunger 13 overcomes the pressing force of the spring 16 and moves the valve body 3 in the valve opening direction. The amount of movement of the valve body 3 is such that the valve body 3 moves to a position balanced with the load obtained by multiplying the composite spring constant K of the spring 16 and the spring 15 by the amount of movement L. As a result, the valve seat 27 opens, and the high-pressure refrigerant flows through the valve orifice 6 to the low-pressure refrigerant flow path 5. That is, since the balance position of the attraction force and the load by K×L changes continuously in accordance with the current value, the opening area formed by the valve seat 27 and the conical portion of the valve body 3 can be continuously varied.

Therefore, the flow rate of the refrigerant can be controlled by controlling the current value applied to the electromagnetic coil 1. The current applied to the electromagnetic coil 1 is controlled using a pulse width modulation method (
P, W, M method) is optimal because it can minimize the flow rate hysteresis.

When the application to the electromagnetic coil 1 is stopped, the conical portion of the valve body 3 is instantly pressed against the valve seat 27 by the pressing force of the spring 16, so that the flow of the refrigerant is instantly stopped. Further, in the conventional valve structure, leakage occurred from the gap, but in this embodiment, the packing seals the leakage from the gap, so the refrigerant sealing function is high. According to the present invention, since the non-magnetic spacer 12 is provided on the attractor 10, it is possible to prevent the attractor 10 and the plunger 13 from coming into close contact even if the current value exceeds IMAX.

Regarding the control accuracy of the flow rate, as shown in FIG. Since the high-pressure refrigerant pressure and the low-pressure refrigerant pressure act equally on the valve body 3 in opposite directions in the axial direction, the load that the valve body 3 receives from these pressures becomes zero.

Further, the half apex angle θ of the conical portion of the valve body 3 in the valve closing direction, which occurs when the refrigerant flows out of the valve orifice 6, reaches a maximum at 45°, and reaches a maximum at 19°.
becomes 0. (An experimental example) Therefore, in this example, this θ
By setting the angle to 15° to 30°, the axial thrust is also set to 0.

Therefore, since the force that drives the valve body 3 is determined by the relationship between the spring force and the suction force, and these forces are stable, the position control (valve stroke control) of the valve body can be performed accurately. The accuracy of the flow rate is good because it can be done.

FIG. 4 shows the suction force characteristics of the first embodiment, with the horizontal axis representing the gap (distance between the opposing surfaces of the attractor 10 and the plunger 13) and the vertical axis representing the suction force. The solid line in the figure is the current IMIN
..., shows the suction force corresponding to IMAX, and the dotted line and the dashed-dotted line show the pressing force of the spring.

[0055] Current value applied to coil 1 IMIN...I
At MAX, the intersection point P1 of the suction force and the spring pressing force...
P4 is the displacement amount by which the plunger 13 moves in the direction of the attractor 10.

FIG. 5 shows this relationship, with the horizontal axis representing the current values IMIN...IMAX applied to the coil 1 and the vertical axis representing the valve stroke L (plunger displacement = air gap). This shows that the valve stroke can be controlled with

However, the intersections P1...P4 in FIG. 4 change to the intersections PB1...PB4 or PC1...PC3 depending on the size of the gap (errors caused by component dimensions and assembly accuracy), so the single point shown in FIG. The valve stroke characteristics are shown by the chain line.

[0058] Therefore, in order to obtain a flow rate of refrigerant corresponding to the applied current value, it is necessary to have precision that allows the size of this gap to be constantly adjusted to a constant value.

Therefore, in this embodiment, the P1 point is determined by using the jig 28 shown in FIG. It is configured so that it can be adjusted. According to such a state, the control accuracy of the flow rate is further improved.

FIG. 7 shows a second embodiment of the present invention.
This is a modification of the valve body structure, which is the main part.

In this embodiment, a conical valve seat 27 is provided in the valve orifice 6 disposed in the valve body 2.
The valve body 3 is made up of a flat disk portion seated on the valve body 3 and a cylindrical shaft portion fitted into the sliding hole 9. By adopting such a configuration, it becomes possible to expand the flow rate range of the first embodiment with the same valve stroke.

In other words, in the first embodiment, the half apex angle of the conical portion is set to an acute angle in order to eliminate the axial thrust acting on the valve body 3, so although the opening area with respect to the valve stroke is small, the axial thrust can also be eliminated by setting the half-vertex angle to 60° or more. Therefore, in this embodiment, a flat plate shape is used that has the maximum opening area that can be obtained with the same valve stroke and the same valve orifice diameter.

[0063] Therefore, by simply changing the valve body structure, approximately 2
The flow range is doubled.

FIG. 8 shows a third embodiment of the present invention, and is a modification of the opposing surfaces of the plunger and the attractor, which are the main parts thereof.

The purpose of this embodiment is to expand the valve stroke range and the flow rate control range by changing the attractor characteristics.

FIG. 9 shows the attraction force characteristics of this embodiment. The suction force has an inflection point with respect to the gap, and after the inflection point, the suction force with respect to the gap gradually changes.

That is, since the suction force obtained in the first embodiment is obtained in a large gap, the amount of movement of the plunger can be increased, and therefore the valve stroke is expanded.

FIG. 10 shows the valve stroke L with respect to current, where the dotted line is the valve stroke of the first embodiment, and the solid line is the valve stroke of the present embodiment.

From the above, it is possible to expand the flow rate range by approximately 1.5 to 2 times by simply changing the shapes of the opposing surfaces of the plunger and the attractor.

In order to eliminate frictional force (sliding resistance) between the spring 15 and each member receiving the spring 16 and reduce flow hysteresis, the springs are fixed at both ends to prevent displacement of the springs in the portions that come into contact with each member. The bottom of the mask is wrapped in multiple layers.

Since the packing 7 has a diameter slightly smaller than the diameter of the cylindrical shaft in order to seal the outer periphery of the cylindrical shaft of the valve body 3, a restraining force is generated. Since this binding force tends to increase as the high pressure increases, the binding force of the packing affects the accuracy of flow rate control, albeit slightly, under high pressure.

Therefore, a labyrinth seal structure was adopted in order to sufficiently reduce the pressure between the sliding hole 9 and the cylindrical shaft and reduce the pressure applied to the packing.

At the same time, the refrigerant sealing function can also be improved.

[0074]

[Effects of the Invention] As explained above, the present invention provides a flow rate control valve that can arbitrarily set the flow rate of refrigerant in proportion to an electrical input signal and is suitable for performing a wide range of control regardless of the type of refrigerant. It is possible.

[Brief explanation of drawings]

FIG. 1 is a sectional view of a first embodiment of the flow control valve of the present invention;

[
FIG. 2 is a sectional view of the valve body portion of the first embodiment,

[Figure 3] A curve diagram showing the dynamic characteristics of the valve body,

FIG. 4 is a curve diagram showing the relationship between the air gap and magnetic force (attractive force) in the first embodiment;

FIG. 5 is a curve diagram showing the relationship between current and valve stroke in FIG. 4;

[Fig. 6] A sectional view showing a method for setting the usage range of the suction force shown in Fig. 4;

FIG. 7 is a sectional view of a valve body portion of a second embodiment of the present invention;

[
FIG. 8 is a cross-sectional view of the opposing surfaces of the plunger and the attractor according to the third embodiment of the present invention;

FIG. 9 is a curve diagram showing the relationship between the air gap and magnetic force (attractive force) in the third embodiment;

FIG. 10 is a curve diagram showing the relationship between current and valve stroke according to FIG. 9;

FIG. 11 is a sectional view showing an example of a conventional flow control valve;

FIG. 12 is a sectional view showing another example of a conventional flow control valve;

[Figure 1
3] A sectional view showing still another example of a conventional flow control valve.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1...Electromagnetic mechanism, 2...Valve body, 3...Valve body, 4...High pressure side circulation hole, 5...Low pressure side circulation hole, 6...Valve orifice, 9...Sliding hole, 10...Suction element, 11...Rod, 12 ...Spacer, 13
...Plunger, 14...Spring pin.

Claims (7)

[Claims] Claim 1: The valve body has a through hole including a low pressure side flow hole that intersects with a high pressure side flow hole, one side is formed with a sliding hole for the valve body, and the other side is provided with a valve orifice, and the sliding hole is formed in the valve body. A valve body is provided that is movably inserted into the hole in the axial direction and moves into and out of contact with the valve orifice, and the valve body is driven by an electromagnetic mechanism provided in a housing fixed to the valve body, and the electromagnetic force is applied to the electromagnetic mechanism. In order to adjust the range of use, a cylindrical shaft-shaped rod made of non-magnetic material and a spring pin are arranged inside the plunger.
The flow control valve further comprises a hollow thin disk spacer made of a non-magnetic material for regulating the amount of movement of the valve body, and arranged between opposing surfaces of the plunger and the attractor. 2. The shape of the valve body consists of a cylindrical part with the same diameter as the valve orifice diameter and a conical part with an acute angle of 30° to 60°. The flow control valve according to claim 1, characterized in that the effect of refrigerant pressure is eliminated. 3. A valve seat is disposed around the valve orifice portion of the valve body, and the valve body has a shape having a cylindrical portion having the same diameter as the valve orifice diameter and a flat portion that seats on the valve seat. The flow control valve according to claim 1, characterized in that: 4. In order to seal the refrigerant flowing from the high-pressure side through the cylindrical portion of the valve body and from the pressure equalization hole to the low-pressure side, a labyrinth seal structure is formed by the sliding hole and the cylindrical portion, and 2. The flow control valve according to claim 1, further comprising a packing seal structure on the outer periphery of the cylindrical portion. 5. The flow control valve according to claim 1, wherein opposing surfaces of the plunger and the suction element have a planar shape. 6. The flow control valve according to claim 1, wherein one of the facing surfaces of the plunger and the suction element is convex and the other is concave. 7. The flow rate control valve according to claim 1, wherein the end windings on both end faces of the spring disposed on the suction element and the valve body are completely tightly wound multiple times.