JPH11304269A - Refrigerating cycle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To satisfactorily suppress the hunting phenomenon caused by the flush vapor of a refrigerant flowing in a low pressure side expansion valve without adding any heat exchanger, etc. SOLUTION: In a two-stage compression refrigerating cycle in which an expanding process is performed in two stages by means of a high pressure side expansion valve 4 and a low pressure side expansion valve 6 and, at the same time, a gas-liquid separator 5 is provided between the high and low pressure side expanding means 4 and 6, the low pressure side expansion valve 6 is constituted of a constant differential pressure regulating valve which roughly fixes the difference between the intermediate pressure on the separator 5 side and the low pressure on an evaporator 7 side. Since the differential pressure regulating valve controls the differential pressure between the pressures before and after the valve, the pressure after the valve does not change from the pressure before the valve even when the density of a flowing-in refrigerant changes. Consequently, the occurrence of hunting can be prevented even when flush vapor is mixed in the flowing-in refrigerant, because the cycle balance does not change.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a two-stage compression refrigeration cycle suitable for use as an air conditioner for automobiles or an air conditioner for home use.

[0002]

2. Description of the Related Art Compressors, radiators, expansion valves (expansion means),
As a conventional technique for improving a coefficient of performance (COP) in a vapor compression refrigeration cycle including an evaporator, FIG.
16 and the two-stage compression two-stage expansion refrigeration cycle shown in FIG.
The economizer cycle shown in FIG. 8 is used.

[0003] The former two-stage compression two-stage expansion refrigeration cycle has two compressors 101, a low stage (low pressure side) and a high stage (high pressure side).
The refrigerant that has exited the high-stage compressor 102 is radiated and condensed by the radiator (condenser) 2, decompressed to an intermediate pressure by the high-pressure side expansion valve 4, and sent to the gas-liquid separator 5. . In the gas-liquid separator 5, the intermediate-pressure refrigerant inside the gas-liquid separator 5 and the low-stage side (low-pressure side)
The gas discharged from the compressor 101 is mixed. At this time, a part of the liquid refrigerant in the gas-liquid separator 5 evaporates and cools the discharge gas to the saturation temperature.

In the gas-liquid separator 5, the refrigerant is separated into saturated vapor and saturated liquid. The saturated liquid is reduced in pressure from the intermediate pressure to the evaporation pressure (low pressure) by the low pressure side expansion valve 6 and sent to the evaporator 7.
The saturated steam is absorbed by the high-stage compressor 102 and is compressed to a high pressure. In this two-stage compression two-stage expansion refrigeration cycle, the refrigerant changes in state as shown in the Mollier diagram in FIG. 16 and the enthalpy difference in the evaporator 7 can be increased. Can be bigger.

On the other hand, the latter economizer cycle is shown in FIG.
As shown in FIG. 7, the arrangement of the radiator 2, the evaporator 7, and the high-pressure side and low-pressure side expansion valves 4 and 6 is the same as that of the two-stage compression two-stage expansion cycle. The suction port 102a of the high-stage compressor 102 is communicated with the pipe 123, and the communication pipe 12 from the gas-liquid separator 5 is connected to an intermediate portion 124 of the pipe 123, so that the refrigerant gas discharged from the low-stage compressor 101 A mixture of air and saturated vapor from the gas-liquid separator 5 is sucked into the high-stage compressor 102 and compressed. Thereby, similarly to the two-stage compression two-stage expansion cycle, the COP can be increased as compared with the one-stage compression cycle.

In the present specification, the two-stage compression two-stage expansion cycle and the economizer cycle are referred to as a two-stage compression refrigeration cycle.

[0007]

In the two-stage compression refrigeration cycle described above, the liquid refrigerant saturated in the gas-liquid separator 5 is reduced in pressure from the intermediate pressure to the evaporation pressure (low pressure) by the low pressure side expansion valve 6. However, flash vapor may be generated in the saturated liquid refrigerant for the following reasons, and this flash vapor intermittently and unstablely flows into the throttle passage of the low-pressure side expansion valve 6, and the flow resistance between the gas and the liquid is reduced. Is greatly changed, causing problems such as hunting of the expansion valve (cycle).

A first cause is that the pipe 125 from the gas-liquid separator 5 to the low-pressure side expansion valve 6 is long,
When the flow rate of the refrigerant is large, the refrigerant pressure is reduced due to the pressure loss of the refrigerant pipe 125 from the gas-liquid separator 5 to the low-pressure side expansion valve 6, and flash vapor is generated in the saturated liquid refrigerant. The second cause is that, during cooling operation with a high outside air temperature or when the engine room becomes hot as in an air conditioner for automobiles and the cycle piping is exposed to a high-temperature atmosphere, the piping temperature is lower than the saturated liquid temperature. Since the ambient temperature around 125 rises, the saturated liquid absorbs heat from the surroundings to generate flash vapor.

Therefore, it is conceivable to shorten the length of the pipe 125 from the gas-liquid separator 5 to the low-pressure side expansion valve 6 in order to suppress the above-mentioned generation of flash steam. Piping 1 from expansion valve 6 to evaporator 7
As a result, the amount of heat exchange (the amount of heat absorbed by the refrigerant) between the low-temperature refrigerant in the pipe 126 and the atmosphere around the pipe 126 increases, and the cooling capacity decreases, Dew condensation occurs on the surface of the pipe 126, and the droplets cause problems such as corrosion of the surrounding metal.

[0010] Further, in order to suppress the above-mentioned generation of flash vapor, Japanese Patent Application Laid-Open No. 8-121889 filed by the present applicant has been proposed.
In the publication, it is proposed that a heat exchanger for cooling the refrigerant in the pipe is provided in the pipe 125 from the gas-liquid separator 5 to the low-pressure side expansion valve 6. The additional space for the air conditioner is increased by the additional space. Particularly, in the vehicle air conditioner, since the installation space is strongly restricted, the addition of the heat exchanger deteriorates the mountability on the vehicle.

In addition, since piping for the heat exchanger is required and the piping becomes complicated, the mountability on a vehicle is further deteriorated and the assemblability of the air conditioner is further deteriorated. The present invention has been made in view of the above problems,
It is an object of the present invention to provide a refrigeration cycle that can satisfactorily suppress a hunting phenomenon caused by flash steam in a refrigerant flowing into a low-pressure side expansion means with a simple configuration without adding a heat exchanger or the like.

[0012]

In order to achieve the above object, according to the first to eighth aspects of the present invention, the expansion process is performed in two stages by the high pressure side expansion means (4) and the low pressure side expansion means (6). In a two-stage compression refrigeration cycle in which a gas-liquid separator (5) is provided between the high pressure side expansion means (4) and the low pressure side expansion means (6), the low pressure side expansion means (6) It is characterized by comprising a constant differential pressure valve which makes the difference between the intermediate pressure on the liquid separator (5) side and the low pressure on the evaporator (7) side almost constant.

[0013] According to this, since the constant differential pressure valve controls the differential pressure across the valve, the pressure difference across the valve does not change even if the density of the inflow refrigerant changes. As a result, the cycle balance does not change even if flash steam is mixed into the flowing refrigerant, so that the unstable phenomenon of cycle hunting does not occur. Moreover, unlike the prior art, there is no need to provide a special heat exchanger for cooling the refrigerant flowing from the gas-liquid separator (5) to the low-pressure side expansion means (6). The piping can be simplified, and the mountability and assemblability of the refrigeration cycle can be favorably maintained even in applications where the installation space of a vehicle or the like is strongly restricted.

According to the second aspect of the present invention, the gas refrigerant is supplied from the gas-liquid separator (5) to the compressor (1, 101, 10).
Valve means (13, 19) are provided in the pipe (12) leading between the low-stage compression section (1b, 101) and the high-stage compression section (1c, 102) in 2), and the compressor (1, 101) is provided. , 102)
The valve (13, 19) opens and closes the flow path of the pipe (12) according to the high pressure on the discharge side and the low pressure on the evaporator (7).

According to this, by appropriately opening and closing the valve means (13, 19) in accordance with the high pressure and the low pressure, the gas-liquid separator (5) can be operated even in an environment where the cycle operation conditions vary widely. ) Can be prevented from returning to the compressor and the refrigerant flowing back from the compressor to the gas-liquid separator (5), and a normal operating state of the cycle can be maintained.

The valve means according to the second aspect of the invention can be constituted by a valve (13) which is electrically controlled to open and close according to the high pressure and the low pressure. Further, as described in claim 4, the valve means may be constituted by a valve (19) having a valve body (191) that is mechanically displaced in response to high pressure and low pressure. The valve (13) according to the third aspect of the present invention is more specifically a detection signal of a pressure sensor (15) for detecting a high pressure and a detection signal of a pressure sensor (16) for detecting a low pressure as described in the fifth aspect. May be electrically controlled to open and close by the output of the control device (14).

According to a sixth aspect of the present invention, instead of the pressure sensor (16) for detecting the low pressure, a temperature sensor for detecting the saturation temperature of the low pressure side refrigerant or the temperature of the air blown from the evaporator (7) is detected. A temperature sensor (17) may be used. Further, a temperature sensor (18) for detecting the temperature of the air for cooling the radiator (2) may be used instead of the pressure sensor (15) for detecting the high pressure.

Furthermore, if the low-pressure side expansion means (6) is integrated with the gas-liquid separator (5) as in the invention of claim 8, the piping for assembling the cycle becomes easy, and
The housing (51) of the gas-liquid separator (5) and the housing (61) of the low-pressure side expansion means (6) can be integrally formed. Therefore, the number of parts can be reduced, and the manufacturing cost can be reduced.

Note that the reference numerals in parentheses of the above means indicate the correspondence with specific means described in the embodiment described later.

[0020]

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a block diagram showing a first embodiment of the present invention. (First Embodiment) FIG. 1 shows a configuration of a refrigeration cycle according to a first embodiment, in which reference numeral 1 denotes a two-stage compressor.
a, a low-stage compression section 1b for compressing the gas refrigerant from the suction port 1a, a high-stage compression section 1c for further compressing the gas refrigerant compressed by the low-stage compression section 1b, And a discharge port 1d for discharging the compressed refrigerant gas from the compression section 1c.

An injection port 1e, a gas refrigerant from the injection port 1e and a gas refrigerant compressed by the low-stage compression section 1b are provided between the low-stage compression section 1a and the high-stage compression section 1b. Are provided. Specifically, the following one can be used as such a two-stage compressor 1. For example, a compressor in which two working chambers are provided in one compressor, such as a rolling piston type compressor described in JP-A-7-110167, and a scroll-type compressor described in JP-A-4-321786. The fixed scroll is provided with two injection ports to inject the refrigerant into the working chamber during the compression stroke as described in JP-A-9-1518.
As in the swash plate type compressor described in Japanese Patent Application Laid-Open No. 47-47, a configuration in which compression is performed to an intermediate pressure in a compression chamber on a lower stage side and further from an intermediate pressure to a high pressure in a compression chamber on a higher stage side can be used.

Reference numeral 2 denotes a radiator (condenser) for cooling and condensing a high-temperature and high-pressure gas refrigerant, and a blower fan 3 is provided to promote heat exchange. Reference numeral 4 denotes a high-pressure side expansion valve (high-pressure side expansion means) for reducing the pressure of the refrigerant condensed by the radiator 2 from a high pressure to an intermediate pressure. Reference numeral 5 denotes a gas-liquid separator which separates gas from a gas-liquid mixed refrigerant depressurized to an intermediate pressure. The specific configuration of the gas-liquid separator 5 is as shown in FIGS. 2A and 2B, and has a main body case 51 formed in a cylindrical shape. On the other hand, the refrigerant inlet 53 opens in the tangential direction. Accordingly, the refrigerant condensed by the radiator 2 flows from the refrigerant inlet 53 into the main body case 51 with a circumferential velocity component. At this time, since there is a density difference between the gas-phase refrigerant and the liquid-phase refrigerant, the high-density liquid-phase refrigerant moves in the direction of the inner peripheral surface 52 of the main body case 51 due to centrifugal force. Since the gas-phase refrigerant is little affected by the centrifugal force, it gathers near the central axis inside the main body case 51.

A gas extraction pipe 54 for extracting a gaseous refrigerant is provided above the central axis of the main body case 51. In the main body case 51, an outlet 55 is provided on a surface opposite to the surface on which the gas extraction pipe 54 is provided (that is, a lower surface), and the gas-liquid mixed refrigerant flows out therefrom. Reference numeral 56 denotes a disk-shaped baffle plate, which is maintained at a predetermined distance from the inner peripheral surface 52 of the main body case 51 by a plurality of protrusions 56a provided on the outer peripheral portion thereof.
An annular refrigerant flow path 5 along the inner peripheral surface 52 of the main body case 51
7 are formed. Thereby, the liquid refrigerant collected near the inner peripheral surface 52 of the main body case 51 can be moved downward by gravity and directed to the outlet 55 through the annular refrigerant flow path 57. Therefore, pipe 5
The baffle plate 56 can prevent the liquid refrigerant in the vicinity of the inner peripheral surface 52 from being entrained in the flow of the gas-phase refrigerant flowing out toward 4.

Reference numeral 6 denotes a low pressure side expansion valve (low pressure side expansion means).
The liquid refrigerant flowing from the outlet 55 of the gas-liquid separator 5 is expanded under reduced pressure. The low pressure side expansion valve 6 is a constant differential pressure valve that controls the differential pressure between the intermediate pressure and the low pressure to be constant (see FIG. 3 described later).
It consists of. Reference numeral 7 denotes an evaporator, which performs heat exchange (heat absorption) between the low-temperature gas-liquid two-phase refrigerant depressurized to a low pressure by the low-pressure side expansion valve 6 and conditioned air (fluid to be cooled) to evaporate the refrigerant. Let it. Reference numeral 8 denotes an air-conditioning case that houses the evaporator 7, and the conditioned air (inside air or outside air) sucked from the suction port 9 is sent to the evaporator 7 by the centrifugal fan 10.

An accumulator 11 stores excess refrigerant, separates gas-liquid refrigerant flowing out of the evaporator 7, and sucks gas refrigerant into the compressor 1.
The liquid refrigerant is prevented from being sucked into the compressor 1. 12
Numeral denotes an injection pipe for guiding the gas refrigerant from the gas extraction pipe 54 of the gas-liquid separator 5 to the injection port 1e of the compressor 1.

Next, the low pressure side expansion valve 6 which is a feature of the present invention is described.
Will be described in detail. FIG. 3 shows the low pressure side expansion valve 6.
This is an example of a specific configuration of the constant differential pressure valve constituting
Reference numeral 1 denotes a cylindrical housing made of a metal such as stainless steel or brass, and an inlet 62 communicating with the outlet 55 of the gas-liquid separator 5 is opened at one axial end thereof. An outlet 63 communicating with the inlet side of the evaporator 7 is formed at the other end of the housing 61 in the axial direction.

In the housing 61, there is formed a valve port 64 for communicating a space 62a on the inlet 62 side with a space 63a on the outlet 63 side. The space 63a
A valve body 65 for adjusting the opening degree of the valve port 34 is provided therein, and the valve element 65 is pressed toward the valve port 64 by a coil spring (elastic member) 66 made of metal.
The housing 61 includes a lid 61a having an outlet 63 and a bottom 61b having an inlet 62.
And a cylindrical main body 61c. The bottom 61b and the main body 61c are integrally formed, and the lid 61a is provided with the valve body 65 and the coil spring 66 for the main body 6c.
After being housed in 1c, it is joined to the end face of the main body 61c by fastening means such as welding or screws.

Reference numeral 67 denotes a cylindrical guide skirt integrally formed with the valve body 65, and the valve guide 65 is provided inside the housing 61.
This guides the movement of the object. The cylindrical outer surface side 67a of the guide skirt 67 corresponds to the inner wall 6 of the main body 61c.
By contacting 1d, the movement of the valve body 65 is guided. Further, in the guide skirt 67, in the vicinity of the valve body 65, a plurality of holes 67b forming a refrigerant flow path are formed in a circumferential direction.

Next, the operation of the constant pressure differential valve which constitutes the low pressure side expansion valve 6 will be described.
5, the acting force F1 due to the pressure (intermediate pressure) on the outlet side of the gas-liquid separator 5 acts on the inflow port 62 side, and the valve body 65 is pressed toward the outflow port 63 by this acting force F1. . On the other hand, since the acting force F2 due to the pressure on the inlet side of the evaporator 7 and the elastic force of the coil spring 66 acts on the outlet 63 side, the valve body 65 is pressed toward the inlet 62 by the acting force F2.

Therefore, when the acting force F2 is larger than the acting force F1, the valve body 65 moves so that the opening of the valve port 64 becomes smaller. Conversely, when the acting force F1 is larger than the acting force F2. Then, the valve body 65 moves so as to increase the opening of the valve port 34 (see FIG. 3B). Therefore, the valve element 6
5 stops at a position where the acting force F1 and the acting force F2 are balanced (or a position in contact with the valve port 64).
The opening degree of No. 4 is determined by the elastic force exerted by the coil spring 66 on the valve body. That is, the pressure difference ΔP between the two spaces 62a and 63a corresponds to the elastic force exerted on the valve body 65 by the coil spring 66.

Since the moving amount (lift amount) of the valve body 65 is small, the elastic force exerted on the valve body 65 by the coil spring 66 hardly changes due to the movement of the valve body 65. As a result, the pressure difference ΔP between the two spaces 62a, 63a can be maintained substantially constant. Note that the opening degree of the valve port 64 in the present embodiment refers to the distance t between the valve port 64 and the valve body 65.

FIG. 4 shows the relationship between the flow rate of the constant pressure differential valve constituting the low pressure side expansion valve 6 and the differential pressure ΔP. Since this characteristic is hardly affected by the specific volume of the flowing refrigerant, ΔP is kept constant irrespective of the flash vapor amount of the flowing refrigerant. Therefore, even if flash vapor is generated in the inflow refrigerant, hunting does not occur in the cycle. In addition,
When such a constant differential pressure valve is used as the low-pressure side expansion valve 6, since the degree of superheat of the refrigerant sucked from the evaporator 7 into the compressor 1 cannot be controlled, the refrigerant liquid returns to the compressor 1;
The efficiency of the compressor 1 and the reliability of the compressor 1 may be reduced.
Therefore, an accumulator 11 is provided on the suction side of the compressor 1 to separate the liquid refrigerant and prevent the liquid from returning to the compressor 1.

Since the accumulator 11 is provided,
Excess refrigerant can be stored in the accumulator 11.
Therefore, the capacity of the tank for storing the liquid refrigerant, which is conventionally provided in the gas-liquid separator 5, is eliminated, and the size of the gas-liquid separator 5 is avoided. As a result, in this embodiment, the liquid refrigerant flowing out from the outlet 55 of the gas-liquid separator 5 is mixed with flash vapor in advance, and the Mollier diagram is as shown in FIG. That is, the point of the liquid refrigerant flowing out of the gas-liquid separator 5 is conventionally represented by a point B on the saturated liquid line in FIG.
However, in the present embodiment, the point A is reached due to the mixing of the flash steam.

The transition from the saturated liquid at the point B to the gas-liquid two-phase refrigerant at the point A slightly reduces the cooling capacity, but a simple constant pressure differential valve is used as the low pressure side expansion valve 6. The hunting phenomenon of the cycle can be prevented without adding a heat exchanger or the like simply by using. By the way, in the refrigeration cycle of the first embodiment, conditions for establishing the cycle will be described. As described above, the compressor 1 compresses the low-pressure to intermediate pressure in the low-stage compression section 1b, and mixes the refrigerant compressed to the intermediate pressure with the gas refrigerant from the gas-liquid separator 5 in the intermediate-pressure chamber 1f. Further, a refrigerant of a type in which the mixed refrigerant is compressed to a high pressure in the high-stage compression section 1c is set.

Now, when the flow rate of the injection refrigerant from the injection port 1e is 0, the pressure in the intermediate pressure chamber 1f of the compressor 1 can be calculated by the following equation (1).

[0036]

## EQU1 ## Intermediate pressure chamber pressure Pm = (low-stage compression section suction volume V1 / high-stage compression section suction volume V2) ^k × low pressure Ps where k is a specific heat ratio of the refrigerant (constant pressure specific heat and constant volume specific heat and Ratio). If the pressure in the gas-liquid separator 5 is lower than Pm, the refrigerant flows backward from the compressor 1 to the gas-liquid separator 5 side.
The pressure inside the gas-liquid separator 5 needs to be Pm or higher.

Next, assuming that the high pressure and the low pressure of the cycle are constant, and consider the case where the pressure (intermediate pressure) in the gas-liquid separator 5 is changed. Injection refrigerant flow rate (weight flow rate g
The flow ratio gi / g of i) and the flow rate (weight flow rate g) of the refrigerant passing through the evaporator 7 can be expressed by the following Expression 2 from the suction volume of the compressor 1 and the suction specific volume of the refrigerant.

[0038]

## EQU2 ## gi / g = (high-stage compression unit suction volume V2 / low-stage compression unit suction volume V1) × (low-stage suction specific volume v1 / high-stage suction specific volume v2) -1 From the heat balance of the gas-liquid separator 5, gi / g is a / b (a: difference between the enthalpy of the refrigerant at the outlet of the radiator and the enthalpy of the point A from the Mollier diagram (FIG. 5), b; (The difference in the enthalpy of the refrigerant at the outlet of the radiator), and gi / g determined by the compressor 1 is equal to gi / g determined by the gas-liquid separator 5.

When the intermediate pressure is low, the high-stage suction specific volume v
Since 2 is large, gi / g is small, and the point A in the Mollier diagram (FIG. 5) is located on the side with the larger dryness. On the contrary, when the intermediate pressure increases, the high-stage-side suction specific volume v2 decreases, so that gi / g increases and the point A eventually becomes on the saturated liquid line. When the intermediate pressure is higher than the intermediate pressure, the compressor 1 is more than the amount of flash steam generated in the gas-liquid separator 5.
The high-stage side compression section 1c draws the liquid refrigerant from the gas-liquid separator 5 back to the compressor 1, which causes a decrease in cycle efficiency and a decrease in compressor reliability due to liquid compression. .

FIG. 6 summarizes the above relationship. The higher the high pressure, the closer the point of the radiator 2 exit on the Mollier diagram to the right, so that gi when the point A is on the saturated liquid line is high. / G increases, and a cycle is established up to a high intermediate pressure. If the operating conditions determined by the use environment of the cycle are within the operable range of FIG. 6, the cycle configuration as in the first embodiment may be used. However, in an air conditioner for a vehicle, which is operated from a high external temperature in summer to an extremely low temperature in severe winter, the cycle is operated beyond the operable condition range of FIG. Sometimes.

(Second Embodiment) In the second embodiment, as shown in FIG. 7, a solenoid valve is provided in the injection pipe 12 for sending the refrigerant from the gas-liquid separator 5 to the injection port 1e of the compressor 1. (Valve means) 13 is provided,
By controlling the opening and closing of the solenoid valve 13, the cycle operation in the operable condition range of FIG. 6 is to be maintained.

In FIG. 7, reference numeral 14 denotes a control device for controlling the opening and closing of the solenoid valve 13, which comprises, for example, a microcomputer and its peripheral circuits. The control device 14 receives the detection signals of the pressure sensor 15 for detecting the pressure on the high pressure side of the cycle and the pressure sensor 16 for detecting the pressure on the low pressure side of the cycle, and activates the solenoid valve 13 according to the high pressure and the low pressure. Open and close.

Here, carbon dioxide (CO ₂ ) is used as a refrigerant.
An example applied to a cycle using is described. C
In O ₂ cycles, the critical temperature of CO ₂ is about 31 ° C, typically Freon critical temperature (e.g., R12 in 112 ° C) which is used as a refrigerant because significantly lower than, the radiator 2, CO ₂ cares It only cools (dissipates heat) in the phase and does not condense.

The compressor 1 has a high-stage compression section suction volume V2
And the capacity ratio between the lower stage side compression section suction volume V1 (V2 / V1)
When a prototype having a value of 0.86 was manufactured and evaluated, the operable range was as shown in FIG. The operable range also changes depending on the low pressure. In the example of FIG.
Three cases of 3.6 MPa, 4.1 MPa, and 4.6 MPa are shown.

When the differential pressure between the intermediate pressure and the low pressure is set to 2.3 MPa by the low-stage expansion valve 6 composed of a constant differential pressure valve, the intermediate pressures corresponding to the respective low pressures are shown in FIG. The high pressure and the low pressure are detected by the pressure sensors 15 and 16, and correspond to a map stored in advance in a storage unit (ROM) of the control device 14. If the conditions permit the gas injection operation, the electromagnetic valve 13 is set. The electromagnetic valve 13 is configured to be opened and the solenoid valve 13 is closed unless the condition for the gas injection operation is possible.

As a specific example, in FIG. 8A, the low pressure is 4.6 MPa, the high pressure is 9.8 MPa or more, and the solenoid valve 13 is opened.
Close. In the low-pressure side, in a gas-liquid two-phase state, the temperature and the pressure correspond one-to-one. Therefore, instead of the pressure sensor 16 for detecting the low-pressure pressure, a temperature sensor for detecting the low-pressure refrigerant temperature is arranged in the same place. May be.

According to the second embodiment, the solenoid valve 13 is opened under high operating conditions of high pressure, that is, when the outside air temperature is high and the heat load is large, and the cooling effect can be increased by gas injection. A large cooling capacity can be obtained with respect to the machine speed. On the other hand, when the operating condition is low, ie, when the outside air temperature is low and the heat load is small, the solenoid valve 13
, The cooling effect by gas injection is not increased.

Therefore, the function of adjusting the cooling capacity by opening and closing the solenoid valve 13 is exhibited. Therefore, like an air conditioner for a vehicle, the electromagnetic clutch is turned on / off (ON / OF).
F) In the compressor in which the operation of the compressor 1 is intermittently controlled and the cooling capacity is adjusted, the number of intermittent (ON / OFF) operations of the electromagnetic clutch can be reduced by opening and closing the electromagnetic valve 13, and the blowout due to the intermittent clutch operation. Air temperature fluctuations and intermittent shocks can be reduced, so that the air conditioning feeling of the occupants can be improved.

(Third Embodiment) In the gas-liquid two-phase region on the low pressure side of the refrigeration cycle, the temperature and pressure of the refrigerant correspond one to one, and the low pressure refrigerant temperature and the air blown out of the evaporator 7 There is also a correlation with the temperature of In view of this point, the third embodiment uses a temperature sensor 17 for detecting the temperature of the air blown out of the evaporator 7 instead of the temperature sensor for the low-pressure refrigerant, as shown in FIG.

In a vehicle air conditioner, since this temperature sensor is usually used for controlling the on / off of the electromagnetic clutch, it is not necessary to attach a new temperature sensor, and the existing temperature sensor is used as it is. Utilizing this, opening and closing control of the solenoid valve 13 can be performed. (Fourth embodiment) High-pressure side pressure (high pressure) of a refrigeration cycle
Is determined by the cooling capacity of the radiator 2, and the cooling capacity is substantially determined by the temperature of the air for cooling the radiator 2, so that the high pressure can be estimated from the temperature of the air. Therefore, in the fourth embodiment, focusing on this point, as shown in FIG. 10, instead of the high-pressure side pressure sensor 15 of the third embodiment, the temperature of the air for cooling the radiator 2 (outside air temperature) is changed. Temperature sensor 18 to detect
Is used.

In the vehicle air conditioner, the temperature sensor for detecting the temperature of the air for cooling the radiator 2, that is, the outside air temperature, is as follows:
Since it is used for calculating the target outlet temperature into the vehicle cabin, the opening and closing control of the electromagnetic valve 13 can be performed by using the existing temperature sensor as it is without having to newly install an outside air temperature sensor. (Fifth Embodiment) Focusing on the fact that the solenoid valve 13 may be controlled to open and close according to the high and low pressures of the cycle as described above, the fifth embodiment is directed to the solenoid valve 13 of the second to fourth embodiments. Instead of this, as shown in FIG. 11, a spool valve (valve means) 19 that opens and closes according to the differential pressure between high pressure and low pressure is used.

FIG. 12 shows a specific configuration of the spool valve 19. 191 is a spool (valve element), which is a cylindrical housing 1
It is slidable in the axial direction (up and down direction) in the inside of 92. The spool 191 has two piston portions 191a, 1
91b and a shaft portion 191c connecting the two. 193 is an inlet through which an injection refrigerant (gas refrigerant) flows from the gas-liquid separator 5 via the injection pipe 12, and 194 is an outlet through which the injection refrigerant flows out. The spool 191 is configured to be opened and closed by sliding in the axial direction.

The piston portion 19 at one end of the spool 191
A high pressure on the discharge side of the compressor 1 is introduced into the space 195 (see FIG. 12A) between the first housing 1a and the inner wall at one axial end of the cylindrical housing 192 by the communication hole 192a of the housing 192 and the pipe 196. And high pressure. In a space 197 between the piston portion 191b on the other end of the spool 191 and the inner wall on the other end in the axial direction of the cylindrical housing 192, a low pressure on the suction side of the compressor 1 is provided by a communication hole 192b of the housing 192 and a pipe 198. Has been introduced and the pressure is low.

A force is applied to the spool 191 by the spring 199 to push the spool 191 from the low pressure side to the high pressure side. Therefore, when the relationship of (high pressure−low pressure) × pressure receiving area of piston portions 191a and 191b> spring force is satisfied, spool 191 moves to the low pressure side (downward), and spool valve 19 opens in FIG. It becomes a valve state. Therefore, the inflow port 193 and the outflow port 194 communicate with each other, and the refrigerant is joined from the gas-liquid separator 5 to the compressor 1.

Conversely, when the above relationship is not satisfied (when the height difference is within a predetermined value), the spool 191 moves to the high pressure side (upward) due to the spring force of the spring 199, and the spool valve 19 12 (b) is closed. Therefore, communication between the inflow port 193 and the outflow port 194 is cut off,
The refrigerant injection from the gas-liquid separator 5 to the compressor 1 is stopped.

Here, considering the case where carbon dioxide is used as the refrigerant, the valve opening differential pressure of the low pressure side expansion valve 6 is 2.3 MPa.
a, assuming that the valve opening differential pressure of the spool valve 19 is 5.4 MPa, as shown in FIG. 13, when the low pressure is 3.6 MPa, the intermediate pressure is 5.9 MPa, and when the high pressure is 9.0 MPa or more, the spool valve 19 Opens. When the low pressure is 4.1 MPa, the intermediate pressure is 6.4 MPa, and when the high pressure is 9.5 MPa or more, the spool valve 19 opens.

When the low pressure is 4.6 MPa, the intermediate pressure is 6.9 MPa, the high pressure is 10 MPa or more, and the spool valve 1
9 opens. As described above, when the pressure difference increases to a predetermined value or more, the spool valve 19 opens to prevent the liquid refrigerant from returning from the gas-liquid separator 5 to the compressor 1 via the injection pipe 12. .

(Sixth Embodiment) In the sixth embodiment, FIG.
As shown in FIG. 4, a constant differential pressure valve constituting a low pressure side expansion valve 6 is integrated with a gas-liquid separator 5. By this integration, piping for assembling the cycle becomes easy, and the housing 51 of the gas-liquid separator 5 and the housing 61 of the low-pressure side expansion valve 6 are integrated, so that the number of parts is reduced and the manufacturing cost is reduced. Can be reduced.

(Other Embodiments) The present invention is not limited to a refrigeration cycle for vehicle air conditioning, but can be applied to a refrigeration cycle for other uses such as a house. In the above-described embodiment, the two-stage compressor 1 is provided with the low-stage compression unit 1a and the high-stage compression unit 1b provided in one compressor. As shown in (2), it is a matter of course that a two-stage compressor may be configured by using two independent compressors 101 and 102.

[Brief description of the drawings]

FIG. 1 is a refrigeration cycle diagram showing a first embodiment of the present invention.

FIG. 2A is a cross-sectional view of the gas-liquid separator 5 of FIG.
3B shows an AA cross section. (B) is BB sectional drawing of (a).

3 (a) is a longitudinal sectional view of a closed state of a constant differential pressure valve constituting the low pressure side expansion valve 6 of FIG. 1, and FIG. 3 (b) is a longitudinal sectional view of the same differential valve in an open state.

FIG. 4 is an operation characteristic diagram of a constant differential pressure valve constituting the low pressure side expansion valve 6 of FIG. 3;

FIG. 5 is a Mollier diagram of the refrigeration cycle of FIG. 1;

FIG. 6 is an operation explanatory diagram of the refrigeration cycle of FIG. 1;

FIG. 7 is a refrigeration cycle diagram showing a second embodiment.

FIG. 8 is an explanatory view of an opening / closing operation of a solenoid valve 13 according to a second embodiment.

FIG. 9 is a refrigeration cycle diagram showing a third embodiment.

FIG. 10 is a refrigeration cycle diagram showing a fourth embodiment.

FIG. 11 is a refrigeration cycle diagram showing a fifth embodiment.

FIG. 12A is a longitudinal sectional view of the spool valve 19 of the fifth embodiment in an open state, and shows a BB section of FIG. 12C.
(B) is a longitudinal sectional view of the valve in a closed state, and (c) is an AA sectional view of (a).

FIG. 13 is an explanatory diagram of an opening / closing operation of a spool valve 19 according to a fifth embodiment.

FIG. 14 is a longitudinal sectional view showing an integrated structure of a gas-liquid separator 5 and a low-pressure side expansion valve 6 according to a sixth embodiment.

FIG. 15 is a cycle diagram of a conventional two-stage compression two-stage expansion refrigeration cycle.

FIG. 16 is a Mollier diagram of the refrigeration cycle of FIG.

FIG. 17 is a cycle diagram of a conventional economizer cycle.

FIG. 18 is a Mollier diagram of the economizer cycle of FIG. 17;

[Explanation of symbols]

1, 101, 102: compressor, 2: radiator, 4: high-pressure side expansion valve, 5: gas-liquid separator, 6: low-pressure side expansion valve, 7: evaporator.

Claims (8)

[Claims] 1. A vapor compression refrigeration cycle comprising a compressor (1, 101, 102), a radiator (2), expansion means (4, 6), and an evaporator (7). Means (4) and low-pressure side expansion means (6) in two stages, and a gas-liquid separator (5) between the high-pressure side expansion means (4) and the low-pressure side expansion means (6). A low-stage compression section (1b, 10b) for sucking and compressing the refrigerant at the outlet side of the evaporator (7) in the compressor (1, 101, 102).
1) and a high-stage compression section (1c, 102) for compressing the refrigerant discharged from the low-stage compression section (1b, 101) and the gas refrigerant separated by the gas-liquid separator (5). A two-stage compression refrigeration cycle, wherein the low-pressure-side expansion means (6) is configured so that a difference between an intermediate pressure on the gas-liquid separator (5) side and a low-pressure pressure on the evaporator (7) side is substantially constant. A refrigeration cycle comprising a constant differential pressure valve. 2. A low-pressure side compression section (1) of the compressor (1, 101, 102) for transferring a gas refrigerant from the gas-liquid separator (5).
b, 101) and valve means (13, 19) are provided in the middle of the pipe (12) leading between the high-stage compression section (1c, 102), and the compressor (1, 101, 102) on the discharge side. The refrigeration cycle according to claim 1, wherein the flow path of the pipe (12) is opened and closed by the valve means (13, 19) according to a high pressure and a low pressure on the side of the evaporator (7). . 3. The refrigeration cycle according to claim 2, wherein said valve means comprises a valve (13) electrically controlled to open and close according to said high pressure and said low pressure. 4. The valve according to claim 2, wherein the valve means comprises a valve (19) having a valve body (191) that is mechanically displaced in response to the high pressure and the low pressure. Refrigeration cycle. 5. A pressure sensor (1) for detecting the high pressure.
5) and a pressure sensor (16) for detecting the low pressure
And a control device (14) to which detection signals of the two pressure sensors (16) are input, wherein the valve (13) is electrically opened and closed by an output of the control device (14). The refrigeration cycle according to claim 3, characterized in that: 6. A pressure sensor (1) for detecting the low pressure.
6. The temperature sensor according to claim 5, wherein a temperature sensor for detecting a saturation temperature of the low-pressure side refrigerant or a temperature sensor for detecting a temperature of air blown out from the evaporator is used instead of 6). Refrigeration cycle. 7. A pressure sensor (1) for detecting the high pressure.
The refrigeration cycle according to claim 5, wherein a temperature sensor (18) for detecting a temperature of air for cooling the radiator (2) is used instead of (5). 8. The refrigeration cycle according to claim 1, wherein the low-pressure side expansion means (6) is integrated with the gas-liquid separator (5).