Pressure reducer and refrigerating cycle unit using the same

CROSS REFERENCE TO RELATED APPLICATION

This application is based on and incorporates herein by reference Japanese Patent Application Nos. 2000-105276 filed on Apr. 6, 2000, 2000-189600 filed on Jun. 23, 2000, and 2000-337838 filed on Nov. 6, 2000.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a pressure reducer in a refrigeration cycle unit suitable for use in a vehicle air-conditioner.

2. Description of Related Art

A temperature type pressure reducer has been normally used as a pressure reducer to automatically control the flow rate of refrigerant so that the degree of superheat of refrigerant at the output of an evaporator is maintained at a predetermined value because the width of fluctuations of cycle operating condition is large in a vehicular air-conditioning refrigeration cycle unit. However, the structure of the temperature pressure reducer is complicated and is expensive because it requires a valve driving mechanism which operates corresponding to the degree of superheat of the refrigerant at the output of the evaporator.

Then, there has been proposed a pressure reducer having a simple structure by eliminating the valve driving mechanism in JP-A-11-257802. In this prior art, a pressure reducer having a valve mechanism for changing a restrict diameter corresponding to differential pressure (difference between high pressure and low pressure of the cycle) before and after the pressure reducer is constructed as shown in

FIG. 22

in a refrigeration cycle unit. In the accumulator type refrigeration cycle unit, an accumulator for collecting liquid refrigerant by separating gas and liquid of the refrigerant is disposed between the outlet of the evaporator and the suction side of the compressor.

According to the prior art, the valve mechanism expands the restrict diameter when the circulating flow rate of the cycling refrigerant is balanced with the radiating capability of the condenser and the differential pressure is smaller than a first predetermined value P

1

in running normally for example. Then, the valve mechanism reduces the restrict diameter when the radiating capability of the condenser drops due to the reduction of the cooling air amount and the high pressure increases, thus increasing the differential pressure more than the first predetermined value P

1

in idling. Then, the valve mechanism expands the restrict diameter again when the flow rate of the cycling refrigerant rises remarkably due to the high-speed rotation of the compressor in running at high-speed for example and the high pressure rises further, thereby increasing the differential pressure more than a second predetermined value P

2

.

Thus, the valve mechanism lowers the low pressure by reducing the restrict diameter in idling to assure the cooling capability in idling and expands the restrict diameter in running at high-speed to prevent the high pressure from rising abnormally in the prior art.

However, the actual relationship between the refrigeration cycle operating condition and the differential pressure (difference of high pressure and low pressure in the cycle) before and after the pressure reducer is not determined uniquely as shown in FIG.

22

. For instance, there is a case when the high pressure rises and the differential pressure becomes greater than the second predetermined value P

2

when the radiating capability of the condenser drops extremely even in idling when the outside temperature is high or when the traffic jam occurs in a city and the valve mechanism expands the restrict diameter similarly to the case of running at high-speed. As a result, there arise problems that the low pressure (refrigerant evaporating temperature) rises and the subcooling degree of the refrigerant at the outlet of the condenser reduces, thereby dropping the cooling capability.

A vehicular transmission gear is shifted to low-speed gear and the flow rate of the cycling refrigerant rises remarkably due to the high-speed rotation of the compressor in running an uphill road even in running normally. However, since the car speed is low in running the uphill road, it is often unable to obtain the cooling air amount of the condenser corresponding to the rise of the flow rate of the refrigerant. As a result, there is a case when the high pressure rises and the differential pressure becomes greater than the first predetermined value P

1

as the radiating capability of the condenser becomes insufficient. The valve mechanism reduces the restrict diameter similarly to the case in idling at this time. Thereby, the high pressure rises further, thereby increasing the driving power of the compressor and worsening the efficiency of the cycle.

SUMMARY OF THE INVENTION

In view of the problems described above, an object of the present invention is to provide a pressure reducer having the small and simple structure and capable of controlling the flow rate of refrigerant favorably even when the operating condition fluctuates widely.

In the accumulator type refrigeration cycle unit in which an accumulator for collecting liquid refrigerant by separating the gas and liquid of the refrigerant is disposed between the outlet of the condenser and the intake side of the compressor as disclosed in JP-A-11-257802, saturated gas refrigerant is taken in from the accumulator and is compressed and discharged. Then, the condition (subcooling degree or dryness) of the refrigerant at the outlet of the condenser changes due to the fluctuations of the cycle operating condition. It is effective to maintain the subcooling degree of the refrigerant at the outlet of the condenser in an adequate range (around 7-15° C.) in order to improve the efficiency of the refrigeration cycle.

That is, when the subcooling degree of the refrigerant at the outlet of the condenser becomes excessively large, the driving power of the compressor increases due to the rise of the high pressure. When the subcooling degree of the refrigerant at the outlet of the condenser becomes excessively small in contrary, the difference of enthalpy between the inlet and outlet of the evaporator decreases, thus dropping the capability.

Then, the present invention achieves the above-mentioned object by favorably controlling the flow rate of refrigerant with respect to the wide fluctuations of the driving condition while maintaining the subcooling degree of the refrigerant at the outlet of the condenser in the appropriate range.

According to a first aspect of the present invention, variable restrict means is disposed at the upstream side of flow of the refrigerant. Fixed restrict means is disposed at the downstream side of the variable restrict means, and refrigerant which has passed through the variable restrict means always flows thereto. An intermediate space is provided between said variable restrict means and the fixed restrict means, and passage sectional area of which is larger than that of the fixed restrict means. The length of the intermediate space is larger than a predetermined length required for allowing the refrigerant injected out of the variable restrict means to expand more than a passage sectional area of the fixed restrict means.

The fixed restrict means has the shape of a nozzle or the like. The change of flow rate is large, i.e., a flow rate control gain is large, in the area B where the dryness of refrigerant is small (dryness x<0.1 for example) as indicated by a dot chain line (1) in

FIG. 3

described later.

Then, noticing on this point, the variable restrict means disposed at the upstream side of the flow of refrigerant decompresses the subcool liquid refrigerant at the outlet of the condenser by a predetermined degree to change to the small dryness area, the gas-liquid two phase refrigerant in the small dryness area is flown into the fixed restrict means to decompress again.

Thereby, the refrigerant flow rate control action can be performed in the refrigerant state in which the flow rate control gain is large by the fixed restrict means, so that a large refrigerant flow rate control width D (

FIG. 5

) can be obtained by a small variation width C of the subcooling degree as indicated by (2) in

FIGS. 3 and 5

when the flow rate control action of the fixed restrict means is seen from the relationship with the subcooling degree of the refrigerant at the outlet of the condenser.

Specifically, because the restrict means at the upstream side of the flow of refrigerant is the variable restrict means whose throttle opening can be controlled, an adequate dryness state may be created by the flow rate control action of the fixed restrict means at the downstream side by controlling the throttle opening of the variable restrict means corresponding to the changes of state of the refrigerant at the outlet of the condenser.

Further, the part of the flow of refrigerant where the flow velocity is high and the part thereof where the flow velocity is low may be mixed in the intermediate space by injecting the refrigerant in the small dryness area decompressed by the variable restrict means to the intermediate space where the passage sectional area is larger than that of the fixed restrict means and by expanding the flow of injected refrigerant more than the passage sectional area of the fixed restrict means within the intermediate space. Therefore, the injected flow of refrigerant from the variable restrict means (

14

) can be a flow of relatively uniform flow velocity and this uniform flow of refrigerant may be restricted steadily according to the flow rate characteristic of the fixed restrict means at the downstream side. The flow rate characteristics indicated by (1) in

FIG. 3

may be exhibited steadily by the restricting action of the fixed restrict means.

As a result, the refrigerant flow rate may be controlled in the wide range by the small variation width of the subcooling degree of the refrigerant at the outlet of the condenser even when the refrigeration cycle operating condition fluctuates widely. Therefore, the subcooling degree of the refrigerant at the outlet of the condenser may be kept in an adequately range for improving the efficiency of the cyclic operation, thereby achieving the highly efficient cyclic operation and the assurance of the cooling performance. Further, because it requires no valve driving mechanism which corresponds to the degree of superheat such as temperature type pressure reducer and the small and simple pressure reducer comprising the variable restrict means and the fixed restrict means may be constructed.

According to a second aspect of the present invention the pressure reducer includes bleeding means for allowing the intermediate space to communicate with an upstream side passage of the variable restrict means even when the variable restrict means is closed.

It allows the refrigerant to be flown through the bleeding means even when the variable restrict means is closed, so that it is possible to prevent the variable restrict means from hunting when the flow rate is small while closing the variable restrict means until when the refrigerant flow rate increases to a predetermined flow rate.

According to a third aspect of the present invention, the variable restrict means has a fixed valve seat and a valve body displacing with respect to the fixed valve seat. The valve body displaces in accordance with a pressure difference between at an upstream side and a downstream side thereof.

Thereby, it is possible to keep the pressure difference at a constant value regardless of the fluctuations of the operating condition and to maintain the flow rate control action of the fixed restrict means at the downstream side in a favorable state at all times by changing the subcool liquid refrigerant at the outlet of the condenser to the small dryness area by the variable restrict means.

According to a fourth aspect of the present invention, the pressure reducer includes spring means for urging the valve body toward a valve closing direction against the pressure difference, and the spring force of the spring means is adjustable.

Thereby, the pressure difference may be controlled by setting the spring force of the spring means and the target subcooling degree of the refrigerant at the outlet of the condenser may be readily controlled by controlling the pressure difference. Accordingly, the target subcooling degree may be controlled readily by controlling the spring force of the spring means even when heat exchanging capability is difference due to the change of size of the condenser and the evaporator and when the heat radiating condition of the condenser is changed.

According to a fifth aspect of the present invention, the pressure reducer includes a body member for containing the variable restrict means. The fixed valve seat is assembled to the body member so that its position can be adjusted and the spring force of the spring means is adjusted by adjusting the position of the fixed valve seat.

Thereby, the target subcooling degree may be adjusted readily by adjusting the position of the fixed valve seat with respect to the body member.

According to a sixth aspect of the present invention, the pressure the spring force of the spring means is preset at 3-5 kg/cm

2

.

According to the experiments and study conducted by the inventors, it was found that the subcooling degree of the refrigerant at the outlet of the condenser may be set at the optimum range for improving the efficiency of the cyclic operation and for assuring the cooling performance and that the favorable flow rate control characteristics which allows the refrigerant flow rate to be largely changed by the small variation of the subcooling degree may be obtained by setting the spring preset pressure within that range.

According to a seventh aspect of the present invention, the variable restrict means has a restrict passage formed into a shape such that the refrigerant having contracted at an inlet thereof adheres to an inner wall surface of the intermediate space to be decompressed by tubular friction.

Since the tubular frictional force has the relationship that it is proportional to the square of the flow velocity, it is possible to increase the opening of the variable restrict means by utilizing that the tubular frictional force increases when the flow rate is high. It also allows the action of keeping the pressure difference constant regardless of the fluctuations of flow rate to be enhanced further, thus maintaining the good refrigerant flow rate characteristics (flow rate control gain).

According to an eighth aspect of the present invention, length L

2

of the restrict passage and an equivalent diameter d

2

of the restrict passage satisfy a relation L

2

/d

2

≧5.

According to the study conducted by the inventors, it was found that the operation and effect of the eighth aspect of the present invention can be obtained when the shape of the restrict passage is set so that the above-mentioned ratio becomes L

2

/d

2

>5 in concrete because the decompression effect by the tubular friction in the restrict passage is favorably exhibited.

It is noted that the equivalent diameter means that when the cross sectional shape of the restrict passage is a normal circle, the diameter of the circle is applied as it is and when it is non-circle such as ellipse, it is replaced to a circle of the equal cross sectional area and the diameter of the replaced circle is applied.

According to a ninth aspect of the present invention, it is possible to catch foreign materials within the refrigerant at the upstream side of the variable restrict means and to prevent the small passage section of the pressure reducer from clogging by the foreign materials by disposing a filter at the upstream side of the variable restrict means.

According to a tenth aspect of the present invention, the fixed valve seat is disposed at the upstream side of the valve body and the filtering is assembled in a body with the fixed valve seat.

Thus, the filter may be formed in a body with the fixed valve seat of the variable restrict means, thereby decreasing a number of parts.

According to an eleventh aspect of the present invention, the whole pressure reducer may be constructed as a thin and long cylinder by containing the variable restrict means and the fixed restrict means linearly on a same axial line within a cylindrical body member. Accordingly, the pressure reducer may be disposed readily on the way of cooling pipes even in a very small mounting space such as a vehicular engine room.

According to a twelfth aspect of the invention, a refrigeration cycle unit comprises a compressor for compressing and discharging refrigerant, a condenser for condensing the refrigerant from the compressor, a pressure reducer for decompressing the refrigerant from the condenser, an evaporator for evaporating the refrigerant which has been decompressed by the pressure reducer, and an accumulator for storing the refrigerant from the evaporator. The pressure reducer is composed of the pressure reducer described above.

The invention can exhibit the refrigerant flow rate control action effectively in such accumulator type refrigeration cycle unit.

According to a thirteenth aspect of the present invention, the compressor is driven by a vehicular engine, the condenser is disposed at the region where it is cooled by receiving running wind in running the vehicle and the evaporator cools air blown out to a car room.

Although the state (subcooling degree) of the refrigerant at the outlet of the condenser is inclined to change largely due to the fluctuations of rotational speed of the compressor, to the fluctuations of radiating capability of the condenser caused by the fluctuations of car velocity and to the fluctuations of cooling thermal load of the evaporator in the vehicular accumulator type refrigeration cycle unit, the present invention allows the refrigerant flow rate to be favorably controlled and the subcooling degree of the refrigerant at the outlet of the condenser to be maintained in the adequate range even when the operating conditions fluctuate as described above.

BRIEF DESCRIPTION OF THE DRAWINGS

Additional objects and advantages of the present invention will be more readily apparent from the following detailed description of preferred embodiments thereof when taken together with the accompanying drawings in which:

FIG. 1

is a schematic view showing a refrigeration cycle (first embodiment);

FIG. 2A

is a cross-sectional view showing a pressure reducer (first embodiment);

FIG. 2B

is an enlarged view of showing a main part of the pressure reducer (first embodiment);

FIG. 3

is a characteristic chart of refrigerant flow rate for explaining an operation of the refrigeration cycle (first embodiment);

FIG. 4

is a Mollier chart for explaining the operation of the refrigeration cycle (first embodiment);

FIG. 5

is a characteristic chart of the refrigerant flow rate for explaining the operation of the refrigeration cycle (first embodiment);

FIG. 6

is a characteristic chart of the refrigerant flow rate showing changes of subcooling degree in controlling spring preset pressure (first embodiment);

FIG. 7

is a graph of experimental data showing a relationship between the spring preset pressure and the subcooling degree (first embodiment);

FIG. 8

is a graph of experimental data showing a relationship between the spring preset pressure and the flow rate control gain (first embodiment);

FIG. 9

is a graph for explaining a definition of the flow rate control gain in

FIG. 8

(first embodiment);

FIG. 10

is a characteristic chart of the refrigerant flow rate showing changes of subcooling degree in accordance with the spring preset pressure (first embodiment);

FIG. 11

is a characteristic chart showing a relation ship between a spring lift and the refrigerant flow rate for explaining the operation of the refrigeration cycle (first embodiment);

FIG. 12

is a cross-sectional view showing a pressure reducer (second embodiment);

FIG. 13

is a cross-sectional view showing a pressure reducer (third embodiment);

FIG. 14

is a cross-sectional view showing a main part of a pressure reducer (fourth embodiment);

FIG. 15

is a characteristic chart showing a relationship between a refrigerant flow rate and differential pressure before and after a variable restrict valve (fourth embodiment);

FIG. 16

is a characteristic chart showing a relationship between subcooling degree and the refrigerant flow rate at the inlet of the valve (fourth embodiment);

FIGS. 17A and 17B

are cross-sectional views for explaining pressure-reducing action of the variable restrict valve (fourth embodiment);

FIGS. 18A and 18B

are diagrams for explaining the relationship of force balance acting on the variable restrict valve (fourth embodiment);

FIG. 19

is a graph of experimental data showing a relationship between subcooling degree and the refrigerant flow rate at the inlet of the valve (fourth embodiment);

FIGS. 20A and 20B

are cross-sectional view of an evaluating item used for evaluation of the refrigerant flow rate characteristics of the pressure reducer (fourth embodiment);

FIGS. 21A and 21B

are graphs of experimental data showing the evaluation result of the refrigerant flow rate characteristics in the evaluating item in

FIGS. 20A and 20B

(fourth embodiment), and

FIG. 22

is a characteristic chart showing a relationship between differential pressure before and after a pressure reducer and a restrict diameter (prior art).

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

(First Embodiment)

FIG. 1

shows a refrigeration cycle of vehicular air-conditioning system according to a first embodiment, wherein a compressor

1

is driven by a vehicular engine not shown via an electromagnetic clutch

2

. High pressure gas refrigerant discharged out of the compressor

1

flows into a condenser

3

and is cooled and condensed through heat exchange with the outside air. It is noted that the condenser

3

is disposed at the region, e.g., the front most part within the vehicular engine room in concrete, where it is cooled by receiving running wind in running the vehicle. It is cooled by the running wind and by air blown by a condenser cooling fan.

Then, the liquid refrigerant condensed by the condenser

3

is then decompressed by an pressure reducer

4

to low pressure and is put into the misty gas-liquid two phase state. The pressure reducer

4

is what a plurality of steps of throttle means are disposed in the direction of flow of the refrigerant and its detail will be described later. The low-pressure refrigerant which has passed through the pressure reducer

4

evaporates in an evaporator

5

by absorbing heat from air blown from an air-conditioning fan

6

.

The evaporator

5

is disposed within an air-conditioning case

7

and cold air which has been cooled by the evaporator

5

and whose temperature has been controlled by a heater core section not shown is then blown out to a car room as is well known. The gas refrigerant which has passed through the evaporator

5

is suctioned to the compressor

1

after when an accumulator

8

separates the gas from the liquid.

The accumulator

8

separates the liquid refrigerant from the refrigerant at the outlet of the evaporator

5

to collect the liquid refrigerant, and allows the compressor

1

to suction the gas refrigerant and oil melting in the liquid refrigerant collected at the bottom side of a tank.

FIG. 2A

illustrates the structure of the pressure reducer

4

of the first embodiment, wherein a refrigerant pipe

10

is disposed between the outlet side of the condenser

3

and the inlet side of the evaporator

5

and is usually formed of metal such as aluminum. A body

11

of the pressure reducer

4

is built inside of the refrigerant pipe

10

. This body

11

is molded approximately in a cylindrical shape by resin for example and is positioned by a stopper

12

within the refrigerant pipe

10

.

Sealing O-rings

13

are held in concave grooves

11

a

at the outer peripheral surface of the body

11

. The body

11

is held at the position determined by the stopper section

12

by press-fitting the O-rings

13

into the inner wall surface of the refrigerant pipe

10

.

The pressure reducer

4

is constructed within the body member

11

and includes the following three elements. The first one is a variable restrict valve

14

disposed at the upstream side of the flowing direction A of the refrigerant, the second one is a fixed restrictor

15

disposed at the downstream side of the variable restrict valve

14

, and the third one is an intermediate space (approach space)

16

provided between the variable restrict valve

14

and the fixed throttle

15

.

The variable restrict valve

14

has a fixed valve seat

17

, a valve body

18

which is displaceable with respect to the fixed valve seat

17

and a coil spring

19

for effecting spring force to the valve body

18

in the valve closing direction. The fixed valve seat

17

and the valve body

18

are molded by resin and the coil spring

19

is made of metallic spring member.

The fixed valve seat

17

has a disc portion

17

a

and a cylindrical portion

17

b

formed in a body with the center part of the disc portion

17

a

. A small bleed port

17

c

is formed at the center of the cylindrical portion

17

b

. This bleed port

17

c

composes communicating means for always communicating the intermediate space

16

with an upstream passage

20

of the variable restrict valve

14

with a small opening even when the variable restrict valve

14

is closed as shown in FIG.

2

A. The diameter d

1

of the bleed port

17

c

is as small as φ1.0 mm for example.

The disc portion

17

a

has bypass ports

17

d

around the cylindrical portion

17

b

. The bypass ports

17

d

is divided into a plurality of ports around the cylindrical portion

17

b

in the shapes of arc, circle and the like. The plurality of bypass ports

17

d

allow an enough amount of refrigerant to flow by bypassing the bleed port

17

c

when the variable restrict valve

14

is opened (see FIG.

2

B). The total opening cross sectional area of the plurality of bypassing ports

17

d

is set to be as large as several times or more of the opening cross sectional area of the bleed port

17

c.

A thread

17

e

is created at the outer peripheral surface of the disc portion

17

a

so as to fasten and fix the disc portion

17

a

to the inner peripheral surface of the upstream side end of the body

11

. Here, the disc portion

17

a

may be mechanically fixed to the body

11

by using other fixing means instead of fastening and fixing by the thread

17

e.

The valve body

18

is a cylinder wherein a restrict passage

18

a

formed of a circular hole of small diameter is formed at the center thereof. The diameter d

2

of the restrict passage

18

a

is greater than the diameter d

1

of the bleed port

17

c

and is around φ1.8 mm for example. An inclined concave face (upstream end)

18

b

which press-contacts with an edge inclined face

17

f

of the cylindrical portion

17

b

is formed at the upstream side end of the valve body

18

.

Accordingly, the opening area of the inlet section of the restrict passage

18

a

may be controlled by changing the gap between the edge inclined face

17

f

of the cylindrical portion

17

b

and the inclined concave face

18

b

of the upstream side end of the valve body

18

. An enlarged opening portion

18

c

whose opening cross sectional area is enlarged gradually is formed at the downstream side end of the restrict passage

18

a

. The enlarged opening portion

18

c

reduces a sudden enlargement loss of flow of the refrigerant flown out of the outlet section of the restrict passage

18

a.

One end of the coil spring

19

abuts against the downstream side end face of the valve body

18

and the other end is supported to a stepped face

11

b

formed at the inner peripheral face of the body

11

. It is noted that spring force of the coil spring

19

may be set by adjusting the fastening position of the fixed valve seat

17

to the body

11

. That is, the spring force of the coil spring

19

may be set by adjusting the position of the axial direction of the valve body

18

by adjusting the fastening position of the fixed valve seat

17

by the thread

17

e

of the disc portion

17

a.

Since the pressure difference upstream and downstream of the valve body

18

acts on the valve body

18

as force in the valve opening direction and the spring force of the coil spring

19

acts on the valve body

18

as force in the valve closing direction, the valve body

18

is displaced in the axial direction to control the opening area of the inlet part of the restrict passage

18

a

so that the pressure difference is maintained at a predetermined value determined by the spring force of the coil spring

19

. That is, the variable restrict valve

14

works as a constant differential pressure valve and

FIG. 2B

shows a state in which the valve body

18

is displaced to the side of the coil spring

19

, thereby opening the valve.

The fixed restrictor

15

is formed at the most downstream end of the body

11

in the shape of a nozzle having a smooth passage contracting shape whose cross section is circular arc. Although the case of forming the fixed restrictor

15

directly at the most downstream end of the body

11

is shown in the present embodiment, the fixed restrictor

15

may be made of metal or the like separately from the body member

11

and then be combined in a body with the body

11

by the most downstream end by means of insert molding or the like. The diameter d

3

of the smallest section of the fixed restrictor

15

is set to be equal with the diameter d

2

of the restrict passage

18

a

of the valve body

18

(φ1.8 mm for example) in the present embodiment.

The intermediate space

16

causes the fixed restrictor

15

to exhibit its original restricting action by the flow rate characteristics by equalizing the flow velocity of the refrigerant by mixing the part of exhaust flow of the refrigerant whose flow velocity is high and the part whose flow velocity is low by enlarging the flow of refrigerant exhausted out of the restrict passage

18

a

of the variable restrict valve

14

at its upstream side more than the passage cross sectional area of the fixed restrictor

15

at the downstream side thereof.

Here, the diameter d

4

of the intermediate space

16

is fully larger than the diameter d

2

of the restrict passage

18

a

as well as the diameter d

3

of the fixed restrictor

15

(around φ4.8 mm for example) and its length L is set to be longer than the predetermined length required for enlarging the flow of refrigerant exhausted out of the restrict passage

18

a

more than the passage cross sectional area of the fixed restrictor

15

. The length L is around 40 mm in this example.

It is noted that in the structural example shown in

FIG. 2

, the flow of refrigerant exhausted out of the restrict passage

18

a

flows into the fixed restrictor

15

after adhering again to the inner wall face of the intermediate space

16

by the dimension setting described above (diameter d

4

and length L) and by the enlarged opening portion

18

c

at the downstream end of the restrict passage

18

a.

A filter

21

is disposed at the most upstream end of the body

11

. The filter

21

catches foreign materials such as metal cutting dust and the like contained in the refrigerant to prevent the small restrict passage portion in the pressure reducer

4

from clogging. The filter

21

includes a screen

21

a

formed of resin or the like and a ringed resin frame

21

b

for supporting and fixing the screen

21

a

. The frame

21

b

is fixed to the most upstream end of the body

11

by the fitting anchoring structure or the like utilizing the elasticity of the resin.

As shown in

FIG. 2A

, the whole pressure reducer

4

is formed into the thin and long cylindrical shape of small diameter by arranging the filter

21

, the variable restrict valve

14

, the intermediate space

16

and the fixed restrictor

15

linearly on the same axial line along the flow direction A of the refrigerant.

Next, an operation of the first embodiment constructed as described above will be explained. When the compressor

1

is driven by the vehicular engine in

FIG. 1

, the refrigerant circulates within the refrigeration cycle, repeating the cycle of compressing the refrigerant by the compressor

1

, condensing the refrigerant by the condenser

3

, reducing the pressure of the refrigerant by the pressure reducer

4

, evaporating the refrigerant by the evaporator

5

, separating gas and liquid of the refrigerant by the accumulator

8

and suctioning the refrigerant to the compressor

1

.

The operating condition changes widely in the vehicular air-conditioning refrigeration cycle like the fluctuations of discharge ability of the compressor

1

caused by the fluctuations of the speed of the vehicular engine, the fluctuations of radiating capability of the condenser

3

caused by the fluctuations of car speed and the fluctuations of cooling load of the evaporator

5

(the fluctuations of air blowing amount, the fluctuations of temperature and humidity of suctioned air) and others. Accordingly, it is important to adequately control the flow rate of the cycling refrigerant and the subcooling degree of the refrigerant at the outlet of the condenser corresponding to these cycle operating conditions in order to assure the cooling capability and to enhance the efficiency of refrigeration cycle.

FIG. 3

explains the refrigerant flow rate control operation of the pressure reducer

4

according to the first embodiment, wherein the fixed restrictor

15

at the downstream side of the pressure reducer

4

is formed into the shape of a nozzle and its flow rate characteristic is characterized in that the variation of flow rate is large (flow rate control gain is large) in an area B where the dryness of the refrigerant is small (dryness x<0.1 for example) as shown by a dot chain line (i) in FIG.

3

.

In the first embodiment, the variable restrict valve

14

as the stationary differential pressure valve is disposed at the upstream side of the fixed restrictor

15

to reduce the pressure of the refrigerant at the outlet of the condenser

3

by a predetermined value by the pressure reducing action of the variable restrict valve

14

and to flow the refrigerant in the gas and liquid two phase state and in the area where the dryness is small into the fixed restrictor

15

.

This will be explained by using Mollier chart in FIG.

4

. The refrigerant at the outlet of the condenser

3

is in the condition of point “a” and has predetermined subcooling degree SC. When the high-pressure liquid refrigerant having this subcooling degree SC flows into the pressure reducer

4

, it is decompressed by a predetermined value ΔP by the decompressing action of the variable restrict valve

14

at first. Then, the high-pressure refrigerant is shifted to the gas-liquid two phase state (point b) having the small dryness x

1

. Here, because the variable restrict valve

14

plays the function of the stationary differential pressure valve, its decompression width is maintained always at the predetermined value ΔP.

Next, the refrigerant in the gas-liquid two phase state is exhausted from the restrict passage

18

a

of the valve body

18

of the variable restrict valve

14

to the intermediate space

16

and flows into the fixed restrictor

15

through the intermediate space

16

. Here, the intermediate space

16

can make a flow of refrigerant having relatively uniform distribution of flow velocity by mixing the part of the flow of refrigerant exhausted out of the restrict passage

18

a

whose flow velocity is high and the part whose velocity is low.

Accordingly, since the refrigerant having the uniform distribution of flow velocity flows into the fixed restrictor

15

, the flow rate characteristic shown by (i) in

FIG. 3

may be exhibited reliably by the throttle action of the fixed restrictor

15

. When the variable restrict valve

14

at the upstream side and the fixed restrictor

15

at the downstream side are disposed closely, the refrigerant decompressed by the variable restrict valve

14

at the upstream side flows into the fixed restrictor

15

with non-uniform distribution of flow velocity while keeping the influence of the decompression. It invites a result that it is unable to exhibit the refrigerant flow rate characteristics based on the original throttle action of the fixed restrictor

15

.

Thus, the fixed restrictor

15

can perform the refrigerant flow rate control action while changing the subcooling liquid refrigerant at the outlet of the condenser

3

to the small dryness area (in the state in which the flow rate control gain is large). As a result, the flow rate control action of the fixed restrictor

15

turns out as shown by (ii) in

FIGS. 3 and 5

when it is seen from the relationship with the subcooling degree of the refrigerant at the outlet of the condenser. That is, a large refrigerant flow rate control width D (

FIG. 5

) may be obtained by the small variation width C of the subcooling degree.

Accordingly, when the cooling thermal load of the evaporator

5

becomes large and a large refrigerant flow rate is required for example, it is possible to obtained the required refrigerant flow rate just by increasing the subcooling degree of the refrigerant at the outlet of the condenser by a small degree. It suppresses the rise of the compressor power and enhances the efficiency of the cycle operation because it can prevent the subcooling degree from becoming excessive at the time of high load and the high pressure from rising abnormally.

When the cooling thermal load of the evaporator

5

becomes small and only a small refrigerant flow rate is required in contrary, the refrigerant flow rate may be reduced to the level corresponding to the thermal load just by reducing the subcooling degree of the refrigerant at the outlet of the condenser by a small degree. It allows the highly efficient operation of the cycle to be maintained by suppressing the remarkable decrease of the subcooling degree of the refrigerant at the outlet of the condenser even when the load is low and by suppressing the reduction of enthalpy difference between the inlet and the outlet of the evaporator

5

.

It is noted that although the refrigerant flow rate control action of the pressure reducer

4

has been explained above by exemplifying the fluctuations of cooling thermal load of the evaporator

5

, the operating condition fluctuates remarkably in the vehicular air-conditioning refrigeration cycle by the fluctuations of the discharge capability of the compressor

1

due to the fluctuations of engine speed and the fluctuations of radiating capability of the condenser

3

due to the fluctuations of car speed as described above. Accordingly, although the condition of the refrigerant at the outlet of the condenser (subcooling degree or dryness) is apt to change largely along with the fluctuations of such operating condition in the accumulator type refrigeration cycle in

FIG. 1

, it is possible to deal with such fluctuations of operating condition by the first embodiment by largely changing the refrigerant flow rate by changing the subcooling degree by a small degree.

It then becomes possible by the first embodiment to maintain the variation width of the subcooling degree with respect to the fluctuations of the operating condition within a predetermined range within 7 through 15° C., for example, which is efficient in operating the cycle. It thus contributes to the enhancement of the efficiency in operating the cycle.

A broken line (iii) in

FIG. 5

indicates refrigerant flow rate control characteristics in a comparative example using only a capillary tube as a pressure reducer. The capillary tube requires a far large subcooling degree variation width E as compared to the subcooling degree variation width C described above to obtain the refrigerant flow rate control width D described above and hampers the highly efficient operation of the cycle.

Further, as it is understood from the explanation above, the decompression width is always maintained at the predetermined value ΔP because the variable restrict valve

14

works as the stationary differential pressure valve. Accordingly, it is always possible to change the refrigerant flow rate largely by changing the subcooling degree by a small degree even to the wide fluctuations of the operating condition by setting in advance the dryness of the refrigerant at the inlet of the fixed restrictor

15

so that it falls within the dryness small area B in

FIG. 3

in operating in the normal load by selecting this predetermined value ΔP.

When the fixed restrictor as the capillary tube is used as upstream side throttle means of the fixed restrictor

15

, an amount of pressure loss before and after the fixed restrictor changes based on the flow rate characteristics of this upstream side fixed restrictor throttle and the dryness of the refrigerant at the inlet of the downstream side fixed restrictor

15

fluctuates largely, degrading the flow rate characteristics of the downstream side fixed restrictor

15

as indicated by a broken line (iv) in FIG.

3

.

The following merits may be obtained from the first embodiment because the decompression width ΔP of the variable restrict valve

14

may be controlled readily by controlling the spring force of the coil spring

19

by the thread fastening position of the stationary valve seat

17

.

FIG. 6

is a refrigerant flow rate control characteristic chart corresponding to

FIG. 5

, wherein the term “spring preset pressure” is what the spring force of the coil spring

19

is expressed in terms of pressure (unit is kg/cm

2

). (ii) in

FIG. 6

is the refrigerant flow rate control characteristics by the first embodiment in

FIGS. 3 and 5

. (v) is the refrigerant flow rate control characteristics when the screw fastening position of the stationary valve seat

17

is moved to the left side in

FIG. 2

, i.e., to the side in which the spring preset pressure (spring force) of the coil spring

19

is reduced, as compared to the case of the characteristics (ii). (vi) is the refrigerant flow rate control characteristics when the screw fastening position of the stationary valve seat

17

is moved to the right side in

FIG. 2

, i.e., to the side in which the spring preset pressure (spring force) of the coil spring

19

is increased, as compared to the case of the characteristics (ii).

The variable restrict valve

14

is liable to open in case of the refrigerant flow rate control characteristics (v) because the spring preset pressure of the coil spring

19

decreases and the decompression width ΔP of the variable restrict valve

14

decreases due to the characteristics (ii). As a result, the cycle high pressure is balanced with the pressure lower than that of the characteristics (ii) in case of the refrigerant flow rate control characteristic (v), so that the subcooling degree of the refrigerant at the outlet of the condenser becomes a value SC

2

which is smaller than SC

1

in the characteristics (ii).

The restrict valve

14

is hard to open in case of the refrigerant flow rate control characteristics (vi) because the spring preset pressure of the coil spring

19

increases and the decompression width ΔP of the variable restrict valve

14

increases by the characteristics (ii). As a result, the cycle high pressure is balanced with the pressure higher than that of the characteristics (ii), so that the subcooling degree of the refrigerant at the outlet of the condenser becomes a value SC

3

which is greater than SC

1

in the characteristics (ii).

Thus, the subcooling degree of the refrigerant at the outlet of the condenser may be readily controlled by controlling the spring preset pressure of the coil spring

19

of the variable throttle valve

14

, so that the subcooling degree may be readily controlled in the optimum range around 7 through 15° C., for example, for enhancing the efficiency of the cycle operation even when difference of heat exchanging capability occurs due to changes of size of the condenser

3

and the evaporator

5

and difference of radiating amount occurs due to changes of structure in mounting the condenser

3

in the vehicle. It is practically very convenient.

Next, concrete numerical examples of the spring preset pressure of the coil spring

19

of the variable restrict valve

14

will be explained.

FIG. 7

shows experimental data which has been obtained by the inventor of the present invention and which shows the relationship between the spring preset pressure of the spring

19

of the variable throttle valve

14

and the subcooling degree of the refrigerant at the outlet of the condenser. The main experimental conditions in

FIG. 7

are; inlet air temperature of the condenser

3

and the evaporator

5

is 30 through 40° C. and the rotational speed of the compressor

1

is 800 through 3000 rpm.

As it is understood from

FIG. 7

, the subcooling degree of the refrigerant at the outlet of the condenser falls in the range of 7 through 15° C. in the range when the spring preset pressure within the range of 3 through 5 kg/cm

2

.

The subcooling degree range of 7 through 15° C. is the optimum range in operating the refrigeration cycle from the following reasons. That is, the cycle high pressure is liable to rise excessively, thus increasing the compressor power and lowering the cycle efficiency in the state when the subcooling degree exceeds about 15° C. It is not preferable to lower the subcooling degree below about 7° C. because it is liable to reduce the difference of enthalpy between the inlet and the outlet of the evaporator

5

, thus lowering the cooling capability. Thus, the subcooling degree range of 7 through 15° C. is the optimum range from the both aspects of suppressing the compressor power and of assuring the cooling capability.

FIG. 8

shows the relationship between the flow rate control gain of the pressure reducer

4

having the variable restrict valve

14

and the spring preset pressure of the coil spring

19

of the variable restrict valve

14

. Here, the flow rate control gain is the ratio (D/C) of the variation D of the refrigerant flow rate shown in FIG.

9

and the variation C of subcooling degree of the refrigerant at the outlet of the condenser in concrete.

FIG. 10

shows changes of the flow rate control characteristics caused by the spring preset pressure and shows that the variation of flow rate with respect to the changes of the subcooling degree reduces gradually due to the increase of the spring preset pressure. It means that the flow rate control characteristics degrades due to the increase of the spring preset pressure, i.e., that the flow rate control gain reduces.

A broken line C in

FIG. 8

indicates the flow rate control gain of the pressure reducer

4

composed of only the fixed restrictor

15

(having no variable restrict valve

14

). The flow rate control gain is reduced to the level equal to the broken line C when the spring preset pressure exceeds 7 kg/cm

2

. In contrary, it has been found that the flow rate control gain becomes a value (around 15) near the maximum value in the range of spring preset pressure of 3 through 5 kg/cm

2

, exhibiting the favorable flow rate control characteristics.

Next, another feature of the first embodiment will be explained. Since the bleed port

17

c

of small diameter is formed through the cylindrical portion

17

b

of the fixed valve seat

17

of the variable restrict valve

14

, the intermediate space

16

may be communicated always with the upstream passage portion

20

of the variable restrict valve

14

with a small opening by the bleed port

17

c

and the restrict passage

18

a

of the valve body

18

even when the variable restrict valve

14

is closed as shown in FIG.

2

A.

However, when no bleed passage passing through the bleed port

17

c

of small diameter is provided, the variable restrict valve

14

opens even when the flow rate of the refrigerant is small. Then, the variable restrict valve

14

opens in the state when the lift (spring compression degree) of the coil spring

19

is small when the flow rate is small as indicated by a broken line (vii) in

FIG. 11

, the action of the coil spring

19

becomes unstable and the variable restrict valve

14

is liable to cause hunting in the opening/closing operation.

However, since the bleed passage which passes through the bleed port

17

c

is always formed in the first embodiment, the refrigerant flows through the bleed passage passing through the bleed port

17

c

and the closed state of the variable restrict valve

14

is maintained until when the refrigerant increases up to a predetermined amount Q

1

(a flow rate which causes pressure loss corresponding to the predetermined value ΔP described above) as indicated by a solid line (viii) in FIG.

11

. Then, when the refrigerant flow rate exceeds the predetermined amount Q

1

, the lift (spring compression amount) of the coil spring

19

increases suddenly and the variable restrict valve

14

opens. Therefore, it is possible to prevent the hunting of the valve opening operation caused by the small lift of the coil spring

19

.

(Second Embodiment)

In the first embodiment, the bleed port

17

c

of small diameter which always communicates the upstream side and the downstream side of the variable restrict valve

14

has been formed through the cylindrical portion

17

b

of the fixed valve seat

17

of the variable restrict valve

14

. In the second embodiment, a bleed port

18

d

of small diameter is formed through the valve

18

of the variable restrict valve

14

as shown in FIG.

12

. Thereby, the center part of the stationary valve seat

17

becomes a columnar portion

17

b′.

According to the second embodiment, the bleed port

18

d

is provided in parallel with the restrict passage

18

a

of the valve body

18

, so that the bleed port

18

d

always allows the upstream side of the variable restrict valve

14

to communicate with the downstream side thereof even when the variable restrict valve

14

(the valve body

18

) is closed. Accordingly, the bleeding means of the second embodiment can exhibit the same effect with the first embodiment.

(Third Embodiment)

In the first and second embodiments, the frame

21

b

of the filter

21

is fixed to the most upstream end of the body

11

. In the third embodiment, a ringed resin frame

21

b

which protrudes to the upstream side of the flow of the refrigerant is formed by resin in a body with the disc portion

17

a

of the fixed valve seat

17

of the variable restrict valve

14

as shown in

FIG. 13

in the third embodiment so as to support and fix the screen

21

a

by the frame

21

b.

It allows the supporting and fixing portion of the filter

21

to be formed in a body to the fixed valve seat

17

itself and its cost reduction to be achieved by reducing a number of parts.

(Fourth Embodiment)

A fourth embodiment relates to an improvement for increasing the refrigerant flow rate control gain (refrigerant flow rate control width/ subcooling degree) with respect to changes of subcooling degree of the refrigerant at the outlet of the condenser.

FIG. 14

is an enlarged section view of the main part of the pressure reducer

4

, wherein the variable restrict valve

14

works basically as the fixed differential pressure valve which keeps the differential pressure ΔP before and after the variable restrict valve

14

constant as described before. However, the differential pressure ΔP before and after the variable restrict valve

14

increases actually due to the increase of pressure loss at the variable restrict valve

14

part due to the increase of flow rate.

FIG. 15

shows the relationship between the differential pressure ΔP before and after the variable restrict valve

14

and the refrigerant flow rate. The differential pressure ΔP is liable to increase due to the increase of flow rate as indicated by a broken line F in

FIG. 15

in the general construction of the fixed differential pressure valve. Here, the general construction of the fixed differential pressure valve is the orifice type one in

FIG. 18

b

described later. The differential pressure ΔP=high pressure Ph at the upstream side of the valve−pressure of intermediate part Pm. The fourth embodiment aims at the characteristic which keeps the differential pressure ΔP almost constant regardless of the variation of the refrigerant flow rate like a solid line G in FIG.

15

.

When the longitudinal differential pressure ΔP increases due to the increase of the refrigerant flow rate like a broken line F in

FIG. 15

, the high pressure rises and the subcooling degree SC of the refrigerant at the outlet of the condenser increases as it is apparent from Mollier chart in FIG.

4

.

FIG. 16

shows the relationship between the refrigerant flow rate Gr and the subcooling degree SC of the refrigerant at the outlet of the condenser. The higher the flow rate, the larger the subcooling degree SC of the refrigerant at the outlet of the condenser becomes as indicated by a broken line H in

FIG. 16

by the general construction of fixed differential pressure valve.

As a result, the refrigerant flow rate control gain (refrigerant flow rate control width D/subcooling degree variation width E) decreases (degrades) from the characteristics of the broken line H in FIG.

16

.

Then, noticing on the restrict passage

18

a

of the valve body

18

in the variable restrict valve

14

, the fourth embodiment obtains valve characteristics which can keep the differential pressure ΔP before and after the variable restrict valve

14

almost constant regardless of the variation of the refrigerant flow rate as indicated by the characteristic of the solid line G in

FIG. 15

by causing the restrict passage

18

a

to exhibit the decompressing action by its tubular friction similarly to a capillary tube. Thereby, the refrigerant flow rate control gain (refrigerant flow rate control width D/subcooling degree variation width C) is increased like the characteristics of a solid line I in FIG.

16

.

FIG. 17A

shows the pressure reducing action of the variable restrict valve

14

of the fourth embodiment, and

FIG. 17B

shows a comparative example (in the shape of the general orifice type fixed differential pressure valve) of the fourth embodiment. In constructing the variable restrict valve

14

, the restrict passage

18

a

exhibits the pressure-reducing action by its tubular friction similar to the capillary tube when the ratio of length L

2

to diameter d

2

is set as L

2

/d

2

>5, wherein d

2

is the diameter of the restrict passage

18

a

of the valve body

18

and L

2

is the length thereof.

Here, the losses of the pipe system such as an orifice include losses of sudden contraction, tubular friction and sudden expansion. In case of the shape of orifice like the comparative example of

FIG. 17

b

wherein the length L

2

is relatively short as compared to the diameter d

2

of the restrict passage

18

a

, the flow of refrigerant which is contracted suddenly at the inlet portion of the restrict passage

18

a

flows out of the outlet portion of the restrict passage

18

a

to the intermediate space

16

while being separated from the wall surface of the restrict passage

18

a

(in other words, before the flow of refrigerant adheres again to the wall surface). As a result, no tubular frictional force acts because no pressure-reducing effect occurs due to the tubular friction at the restrict passage

18

a.

However, according to the fourth embodiment, it is possible to set the restrict passage

18

a

having length longer than length L

3

which is necessary for the flow of refrigerant separated from the wall surface of the restrict passage

18

a

by suddenly contracting at the inlet portion of the restrict passage

18

a

to adhere again to the wall surface of the passage by setting the ratio of the length L

2

to the diameter d

2

of the restrict passage

18

a

of the valve body

18

as (L

2

/d

2

)>5 as shown in

FIG. 17

a.

Thereby, the restrict passage

18

a

exhibits the pressure-reducing operation by the tubular friction similar to the capillary tube, so that the tubular frictional force acts on the wall surface of the restrict passage

18

a

. Then, according to the fourth embodiment, the relationship of Fs=F

1

+F

2

holds as shown in

FIG. 18

a

, where Fs is the spring force of the coil spring

19

, Fl is force caused by the differential pressure ΔP before and after the valve and F

2

is the tubular frictional force of the restrict passage

18

a

. Meanwhile, no tubular frictional force acts and Fs=F

1

as shown in

FIG. 18

b

in case of the comparative example of the orifice type.

Since the tubular frictional force F

2

is proportional to the square of flow velocity, the tubular frictional force F

2

becomes large when the flow rate is high. Then, the coil spring

19

is pushed in together with the valve body

18

, so that the opening of the inlet portion of the restrict passage

18

a

increases. That is, according to the fourth embodiment in

FIG. 15

, the opening of the inlet portion of the restrict passage

18

a

increases and the differential pressure ΔP reduces due to the increase of the tubular frictional force F

2

as indicated by an arrow a when the flow rate is high.

However, in case of the comparative example of the orifice type, the differential pressure ΔP increases along with the increase of the refrigerant flow rate as shown by a broken line F in

FIG. 15

because the opening of the inlet portion of the restrict passage

18

a

does not increase due to the tubular frictional force F

2

. As a result, according to the fourth embodiment, it is possible to obtain the valve characteristics which can keep the differential pressure ΔP before and after the variable throttle valve

14

almost constant regardless of the increase of the refrigerant flow rate as indicated by a solid characteristic line G in FIG.

15

. It then allows the refrigerant flow rate control gain (refrigerant flow rate control width/subcooling degree variation width) to be increased like a solid characteristic line I in FIG.

16

.

FIG. 19

shows experimental data verifying the effect of improving the refrigerant flow rate control gain according to the fourth embodiment, wherein the flow rate characteristics have been evaluated by fixing the diameter of the restrict passage

18

a

as d

2

=φ1.9 mm and by changing the length L

2

to six lengths of 1, 2, 4, 6, 8 and 10 mm. In terms of the experimental conditions, the refrigerant flow rate was measured by keeping constant the pressure (high-pressure) at the inlet of the variable restrict valve

14

as Ph=1.08 MPa, by keeping constant the pressure (low-pressure) at the outlet of the fixed restrictor

15

as P

1

=0.36 MPa and by using the subcooling degree SC of the refrigerant at the inlet of the variable restrict valve

14

as a parameter. Here, as an experimental object, a single orifice or capillary was used for this verifying experiment.

The refrigerant flow rate was set to be dimensionless by setting the flow rate Gr

sc=0

of the refrigerant of subcooling degree SC=0 at the inlet as 1 and is plotted in the vertical axis as refrigerant flow rate ratio. As it is understood from

FIG. 19

, the refrigerant flow rate may be changed to around 1.5 times by changing the subcooling degree SC=0 to 10° C. when the length L

2

is 10 mm and L

2

/d

2

is greater than 5 (fourth embodiment). However, the refrigerant flow rate changes only 1.25 times or less by changing the subcooling degree SC=0 to 10° C. in the other comparative example (one in which L

2

/d

2

is 4.2 or less).

That is, it can be seen that the refrigerant flow rate control gain may be increased remarkably by setting (L

2

/d

2

)>5 like the fourth embodiment.

FIG. 20A

shows an evaluating item (i) which was actually designed based on the fourth embodiment and

FIG. 20B

shows an evaluating item (ii) as a comparative case. (L

2

/d

2

)=8.3 in the evaluating item (i) and (L

2

/d

2

)=1.4 in the evaluating item (ii).

FIG. 21A

shows changes of the differential pressure ΔP before and after the variable restrict valve

14

with respect to the changes of the refrigerant flow rate. A favorable result of being able to keep the differential pressure ΔP almost in the constant range of around 0.53 through 0.54 MPa in the evaluating item (i) to the changes of the refrigerant flow rate Gr=100 through 200 kg/h. Therefore, it is possible to suppress the variation width of the subcooling degree SC of the refrigerant at the upstream side of the variable restrict valve

14

in the relatively small range of 10 through 15° C. to the changes of the refrigerant flow rate Gr=100 through 200 kg/h by the evaluating item (i) as shown in FIG.

21

B.

However, the variation width of the differential pressure ΔP with respect to the change of the refrigerant flow rate of the evaluating item (ii) becomes far greater than that of the evaluating item (i) as shown in FIG.

21

A. As a result, the variation width of the subcooling degree SC of the refrigerant at the upstream side of the valve expands to the range of 10 through 20° C. with respect to the change of the refrigerant flow rate Gr=100 through 200 kg/h as shown in

FIG. 21B

, thus decreasing (worsening) the refrigerant flow rate control gain.

(Modifications)

It is noted that although the cases of using the fixed restrictor

15

having the shape of a nozzle as the fixed restricting at the downstream side have been explained in the embodiments described above, it is also possible to use an orifice, venturi and the like as the fixed restricting means beside the nozzle.

Further, although the cases of having the bleed ports

17

c

and

18

d

for communicating the passages before and after the variable restrict valve

14

, even when the variable restrict valve

14

is closed, have been explained in the embodiments described above, a vehicular refrigeration cycle unit which automatically stops when the load condition of the cooling thermal load is low, e.g., when the outside air temperature is low, has been put into practical use. The bleed ports

17

c

and

18

d

may be eliminated in such refrigeration cycle unit because the use condition when the refrigerant flow rate becomes small is rare.

Number	Date	Country
2000-105276	Apr 2000	JP
2000-189600	Jun 2000	JP
2000-337838	Nov 2000	JP

Number	Name	Date	Kind
3311131	Zahuranec	Mar 1967	A
3741242	Hansen et al.	Jun 1973	A
4009592	Boerger	Mar 1977	A
4324112	Fujiwara et al.	Apr 1982	A
4375228	Widdowson	Mar 1983	A
5715704	Cholkeri et al.	Feb 1998	A
5966960	Cummings et al.	Oct 1999	A
6182457	Enderle	Feb 2001	B1

Number	Date	Country
0987 505	Mar 2000	EP
1150076	Oct 2001	EP
11-257802	Sep 1999	JP

Pressure reducer and refrigerating cycle unit using the same

Information

Patent Number

Date Filed

Date Issued

Inventors

Original Assignees

Examiners

Agents

CPC

US Classifications

Field of Search

US

International Classifications

Abstract

Description

Claims

Priority Claims (3)

US Referenced Citations (8)

Foreign Referenced Citations (3)