Vapor compression cycle having ejector

A first evaporator is arranged on a downstream side of an ejector, and a second evaporator is connected to a refrigerant suction inlet of the ejector. A refrigerant evaporation temperature of the second evaporator is lower than that of the first evaporator. The first and second evaporators are used to cool a common subject cooling space and are arranged one after the other in a flow direction of air to be cooled.

CROSS REFERENCE TO RELATED APPLICATION

This application is a continuation-in-part from U.S. patent application Ser. No. 11/055,795 filed on Feb. 9, 2005 now U.S. Pat. No. 7,178,359 and is related to Japanese Patent Application No. 2004-41163 filed on Feb. 18, 2004, Japanese Patent Application No. 2004-74892 filed on Mar. 16, 2004, Japanese Patent Application No. 2004-87066 filed on Mar. 24, 2004, Japanese Patent Application No. 2004-290120 filed on Oct. 1, 2004 and Japanese Patent Application No. 2005-37645 filed on Feb. 15, 2005, the contents of which are incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a vapor compression cycle that includes an ejector, which serves as a depressurizing means for depressurizing fluid and which also serves as a momentum transporting pump for transporting the fluid by entraining action of discharged high velocity working fluid, so that such a vapor compression cycle is effectively applicable to, for example, a refrigeration cycle of a vehicle air conditioning and refrigerating system, which performs a passenger compartment cooling air conditioning operation and a refrigerator cooling operation through use of multiple evaporators.

2. Description of Related Art

Japanese Patent No. 1644707 discloses a vapor compression refrigeration cycle ofFIG. 25, in which a portion of a refrigerant passage located downstream of a radiator13branches to two passages51,52. A cooling air conditioner evaporator55for cooling a vehicle passenger compartment is arranged in the passage51, and a refrigerator evaporator56for cooling a refrigerator is arranged in the passage52.

In the refrigeration cycle of Japanese Patent No. 1644707, the flow of the refrigerant is switched between the flow passage51for the passenger compartment cooling air conditioning operation and the flow passage52for the refrigerator cooling operation by switching solenoid valves53,54. In this way, the passenger compartment cooling air conditioning operation, which is performed through use of the cooling air conditioner evaporator55, and the refrigerator cooling operation, which is performed through use of the refrigerator evaporator56, are balanced.

Furthermore, with reference toFIG. 26, Japanese Patent No. 3322263 (corresponding to U.S. Pat. Nos. 6,477,857 and 6,574,987) discloses a vapor compression refrigeration cycle, in which an ejector14is used as a refrigerant depressurizing means and a refrigerant circulating means. In the vapor compression cycle, a first evaporator61is arranged between a refrigerant outlet of the ejector14and a gas-liquid separator63, and a second evaporator62is arranged between a liquid refrigerant outlet of the gas-liquid separator63and a suction inlet14cof the ejector14.

In the vapor compression cycle of Japanese Patent No. 3322263 shown inFIG. 26, the pressure drop, which is induced by the high velocity flow of the refrigerant at the time of expansion of the refrigerant discharged from a nozzle portion14aof the ejector14, is used to draw the gas phase refrigerant, which is discharged from the second evaporator62, through the suction inlet14cof the ejector14. Also, the velocity energy of the refrigerant, which is generated at the time of expansion of the refrigerant in the ejector14, is converted into the pressure energy at a diffuser portion (a pressurizing portion)14bto increase the pressure of the refrigerant, which is discharged from the ejector14. Thus, the pressurized refrigerant is supplied to the compressor12, and thereby the drive force for driving the compressor12can be reduced. Therefore, the operational efficiency of the cycle can be improved.

Furthermore, the two evaporators61,62can be used to absorb heat from and thereby to cool a common space or can be used to absorb heat from and thereby to cool different spaces, respectively.

However, in the case of the refrigeration cycle of Japanese Patent No. 1644707 shown inFIG. 25, the flow passage51, which is used for the passenger compartment cooling air conditioning operation, and the flow passage52, which is used for refrigerator cooling operation, are switched through use of a timer. Thus, during the refrigerator cooling operation, the passenger compartment cooling operation cannot be performed, so that air conditioning feeling of the passenger may be deteriorated. Furthermore, due to a difference in the states of the evaporators55,56after the switching operation, the discharged refrigerant temperature (i.e., the discharged refrigerant pressure) of the compressor12will change significantly. For example, in the case where the thermal load of the currently operated evaporator55,56after the switching operation is relatively large, the compressor12could be operated at the maximum capacity to cause development of the abnormally high pressure in the high pressure side pipe line, which, in turn, could cause stop of the entire operation.

In the case of the vapor compression cycle of Japanese Patent No. 3322263 shown inFIG. 26, the compressor12should receive only the gas phase refrigerant, and the second evaporator62should receive only the liquid state refrigerant. Thus, the gas-liquid separator63, which separates the refrigerant discharged from the ejector14into the gas phase refrigerant and the liquid phase refrigerant, is required. Therefore, the manufacturing costs are disadvantageously increased.

Furthermore, a distributing ratio of the refrigerant to the first evaporator61and to the second evaporator62needs to be determined using the single ejector14while maintaining the refrigerant circulating (gas phase refrigerant drawing) operation of the ejector14. Thus, it is difficult to appropriately adjust the flow rates of the refrigerant of the first and second evaporators61,62.

Furthermore, the two evaporators61,62can be used to absorb heat from and thereby to cool a common space or can be alternatively used to absorb heat from and thereby to cool different spaces, respectively. Also, it is recited that these two evaporators61,62may be used to cool a room.

However, Japanese Patent No. 3322263 does not recite a specific arrangement of the two evaporators61,62for cooling the room by the two evaporators61,62.

Furthermore, another previously proposed refrigeration cycle, which includes a plurality of evaporators, is shown inFIG. 27.FIG. 27is a schematic diagram of the refrigeration cycle, which includes a previously proposed thermostatic expansion valve105. In the refrigeration cycle, a refrigerant circulation passage R is divided into two passages R1, R2at a point located on the downstream side of a radiator102. One evaporator104is provided in the passage R1and is used to perform, for example, passenger compartment cooling air conditioning operation. The other evaporator106is provided in the passage R2and is used to perform, for example, refrigerator cooling operation.

In the case of the refrigeration cycle, which uses the multiple evaporators, such as of a vehicle air conditioning system including a cool box (the refrigerator), the evaporator104for the passenger compartment cooling air conditioning operation and the evaporator106for the refrigerator cooling operation are temperature controlled to the desired evaporation temperatures, respectively, by intermittently opening and closing a solenoid valve107arranged in the refrigerant passage R2for the refrigerator cooling operation to supply the refrigerant to the refrigerant passage R1for the passenger compartment cooling air conditioning operation. Furthermore, the thermostatic expansion valve105and a fixed metering device108are provided as a depressurizing means. InFIG. 27, numeral101indicates a refrigerant compressor, and numeral109indicates a check valve.FIG. 28is a schematic diagram, in which a box type thermostatic expansion valve105is provided in the refrigeration cycle ofFIG. 27.

In a case where an ejector is used in the refrigeration cycle ofFIG. 28, adjustment (e.g., the flow rate adjustment) to correspond with the load changes and effective response to the rapid change in the rotational speed of the compressor are required. To achieve them, Japanese Unexamined Patent Publication No. 2004-44906 (U.S. patent application Publication No. 2004/0007014A1) discloses an ejector, which shows a high efficiency and a high responsibility throughout the entire load range.

However, for example, when the ejector of Japanese Unexamined Patent Publication No. 2004-44906 (U.S. patent application Publication No. 2004/0007014A1) is used in the refrigeration cycle ofFIG. 28, which includes the box type thermostatic expansion valve105, the orientation of the ejector is limited. Thus, there is less freedom in designing of the refrigeration cycle, i.e., the vapor compression cycle having the ejector.

SUMMARY OF THE INVENTION

The present invention addresses the above disadvantages. Thus, it is an objective of the present invention to provide a vapor compression cycle, which includes multiple evaporators and solves or alleviates at least one of the above disadvantages.

To achieve the objective of the present invention, there is provided a vapor compression cycle, which includes a compressor, a radiator, an ejector, a first evaporator, a first branched passage, a first metering means and a second evaporator. The compressor draws and compresses refrigerant. The radiator radiates heat from the compressed high pressure refrigerant discharged from the compressor. The ejector includes a nozzle portion, a gas phase refrigerant suction inlet and a pressurizing portion. The nozzle portion depressurizes and expands the refrigerant on a downstream side of the radiator. Gas phase refrigerant is drawn from the gas phase refrigerant suction inlet by action of a flow of the high velocity refrigerant discharged from the nozzle portion. The pressurizing portion converts a velocity energy of a flow of a mixture of the high velocity refrigerant and the gas phase refrigerant into a pressure energy. The first evaporator evaporates the refrigerant, which is outputted from the ejector, to achieve a refrigeration capacity. A refrigerant outlet of the first evaporator is connected to a suction inlet of the compressor. The first branched passage branches a flow of the refrigerant at a corresponding branching point located between the radiator and the ejector. The first branched passage conducts the branched flow of the refrigerant to the gas phase refrigerant suction inlet of the ejector. The first metering means depressurizes the refrigerant on a downstream side of the radiator. The second evaporator is arranged in the first branched passage. The second evaporator evaporates the refrigerant to achieve a refrigeration capacity.

To achieve the objective of the present invention, there is also provided a vapor compression cycle, which includes a compressor, a radiator, a first metering means, a first evaporator, an ejector, a first branched passage, a second metering means and a second evaporator. The compressor draws and compresses refrigerant. The radiator radiates heat from the compressed high pressure refrigerant discharged from the compressor. The first metering means depressurizes the refrigerant on a downstream side of the radiator. The first evaporator is connected between a refrigerant outlet of the first metering means and a suction inlet of the compressor. The first evaporator evaporates the low pressure refrigerant, which is outputted at least from the first metering means, to achieve a refrigeration capacity. The ejector includes a nozzle portion, a gas phase refrigerant suction inlet and a pressurizing portion. The nozzle portion depressurizes and expands the refrigerant on a downstream side of the radiator. Gas phase refrigerant is drawn from the gas phase refrigerant suction inlet by action of a flow of the high velocity refrigerant discharged from the nozzle portion. The pressurizing portion converts a velocity energy of a flow of a mixture of the high velocity refrigerant and the gas phase refrigerant into a pressure energy. The first branched passage branches a flow of the refrigerant at a corresponding branching point located between the radiator and the first metering means. The first branched passage conducts the branched flow of the refrigerant to the gas phase refrigerant suction inlet of the ejector. The second metering means is arranged in the first branched passage and depressurizes the refrigerant on a downstream side of the radiator. The second evaporator is arranged in the first branched passage on a downstream side of the second metering means. The second evaporator evaporates the refrigerant to achieve a refrigeration capacity.

To achieve the objective of the present invention, there is also provided a vapor compression cycle, which includes a compressor, a radiator, an ejector, a first evaporator and a second evaporator. The compressor draws and compresses refrigerant. The radiator radiates heat from the compressed high pressure refrigerant discharged from the compressor. The ejector includes a nozzle portion, a refrigerant suction inlet, a mixing portion and a pressurizing portion. The nozzle portion depressurizes and expands the refrigerant on a downstream side of the radiator. Refrigerant is drawn from the refrigerant suction inlet by action of a flow of the high velocity refrigerant discharged from the nozzle portion. The high velocity refrigerant discharged from the nozzle portion and the drawn refrigerant supplied from the suction inlet are mixed in the mixing portion. The pressurizing portion converts a velocity energy of a flow of mixed refrigerant, which is mixed through the mixing portion, into a pressure energy. The first evaporator is connected to a downstream side of the ejector. The second evaporator is connected to the suction inlet of the ejector. The first evaporator and the second evaporator are constructed integrally to cool an air flow that is directed to a common subject cooling space.

DETAILED DESCRIPTION OF THE INVENTION

FIRST EMBODIMENT

FIG. 1shows an exemplary case where a vapor compression cycle according to a first embodiment of the present invention is implemented in a vehicle air conditioning and refrigerating system. The vapor compression cycle includes a refrigerant circulation passage11, through with refrigerant is circulated. A compressor12is arranged in the refrigerant circulation passage11. The compressor12draws and compresses the refrigerant supplied thereto.

In the present embodiment, the compressor12is rotated by, for example, a vehicle drive engine (not shown) through a belt or the like. The compressor12is a variable displacement compressor, which can adjust a refrigerant discharge rate through a change in its displacement. The displacement is defined as an amount of refrigerant discharged from the compressor12per rotation of the compressor12. The displacement of the compressor12can be changed by changing an intake volume of the refrigerant in the compressor12.

A swash plate compressor is most commonly used for this purpose and can be used as the variable displacement compressor12. Specifically, in the swash plate compressor, a tilt angle of a swash plate is changed to change a piston stroke and thereby to change the intake volume of the refrigerant. A pressure (control pressure) in a swash plate chamber of the compressor12is changed by a pressure control electromagnetic device12a, which constitutes a displacement control mechanism, so that a tilt angle of the swash plate is externally and electrically controlled.

A radiator13is arranged downstream of the compressor12in a refrigerant flow direction. The radiator13exchanges heat between the high pressure refrigerant, which is discharged from the compressor12, and the external air (external air supplied from outside of the vehicle), which is blown toward the radiator13by a cooling fan (not shown), so that the high pressure refrigerant is cooled.

An ejector14is arranged further downstream of the radiator13in the refrigerant flow direction. The ejector14serves as a depressurizing means for depressurizing the fluid and is formed as a momentum-transporting pump, which performs fluid transportation by entraining action of discharged high velocity working fluid (see JIS Z 8126 Number 2.1.2.3).

The ejector14includes a nozzle portion14aand a suction inlet14c. The nozzle portion14areduces a cross sectional area of the refrigerant passage, which conducts the refrigerant discharged from the radiator13, to isentropically depressurize and expand the high pressure refrigerant. The suction inlet14cis arranged in a space, in which a refrigerant discharge outlet of the nozzle portion14ais located. The suction inlet14cdraws gas phase refrigerant supplied from a second evaporator18. Furthermore, a diffuser portion14b, which serves as a pressurizing portion, is arranged downstream of the nozzle portion14aand the suction inlet14cin the refrigerant flow direction. The diffuser portion14bis formed to progressively increase a cross sectional area of its refrigerant passage toward its downstream end, so that the diffuser portion14bdecelerates the refrigerant flow and increases the refrigerant pressure, i.e., the diffuser portion14bconverts the velocity energy of the refrigerant to the pressure energy.

The refrigerant discharged from the diffuser portion14bof the ejector14is supplied to a first evaporator15. The first evaporator15is arranged in, for example, an air passage of a vehicle passenger compartment air conditioning unit (not shown) to cool the air discharged into the passenger compartment and thereby to cool the passenger compartment.

More specifically, the passenger compartment conditioning air is blown from an electric blower (a first blower)26of the vehicle passenger compartment air conditioning unit toward the first evaporator15. In the first evaporator15, the low pressure refrigerant, which has been depressurized by the ejector14, absorbs heat from the passenger compartment conditioning air and thereby evaporates into gas phase refrigerant, so that the passenger compartment conditioning air is cooled to cool the passenger compartment. The gas phase refrigerant, which has been evaporated in the first evaporator15, is drawn into the compressor12and is re-circulated through the refrigerant circulation passage11.

Furthermore, in the vapor compression cycle of the present embodiment, a first branched passage16is formed. The first branched passage16is branched from a corresponding branching portion of the refrigerant circulation passage11between the radiator13and the ejector14on the downstream side of the radiator13and is then rejoined with the refrigerant circulation passage11at the suction inlet14cof the ejector14.

A first flow rate control valve (a first metering mechanism or a first metering means)17is arranged in the first branched passage16. The first flow rate control valve17controls the flow rate of the refrigerant and depressurizes the refrigerant. A valve opening degree of the first flow rate control valve17can be electrically controlled. The second evaporator18is arranged downstream of the first flow rate control valve17in the refrigerant flow direction.

The second evaporator18is arranged in, for example, a vehicle refrigerator (not shown) to cool an interior of the refrigerator. Internal air of the refrigerator is blown by an electric blower (a second blower)27toward the second evaporator18.

In the present embodiment, the pressure control electromagnetic device12aof the variable displacement compressor12, the first and second blowers26,27and the first flow rate control valve17are electrically controlled by a corresponding control signal outputted from an electronic control unit (ECU).

Next, operation of the present embodiment will be described with reference to the above structure. When the compressor12is driven by the vehicle engine, the refrigerant is compressed in the compressor12, and therefore the high temperature and high pressure refrigerant is discharged from the compressor12in a direction of an arrow A and is supplied to the radiator13. In the radiator13, the high temperature refrigerant is cooled by the external air and is thus condensed. The liquid state refrigerant, which is discharged from the radiator13is divided into a flow of an arrow B passing through the refrigerant circulation passage11and a flow of an arrow C passing through the first branched passage16.

The refrigerant (the arrow C), which passes through the first branched passage16, is depressurized through the first flow rate control valve17and thus becomes the low pressure refrigerant. Then, in the second evaporator18, the low pressure refrigerant absorbs heat from the interior air of the refrigerator, which is blown by the second blower27, so that the refrigerant evaporates. In this way, the second evaporator18cools the interior of the refrigerator.

Here, the refrigerant flow rate in the first branched passage16, i.e., the refrigerant flow rate in the second evaporator18is adjusted by controlling the valve opening degree of the first flow rate control valve17of the first branched passage16through the ECU (a control means)25. Therefore, the cooling capacity of the subject cooling space (specifically, the interior space of the refrigerator), which is cooled by the second evaporator18, is controlled by controlling the valve opening degree of the first flow rate control valve17and a rotational speed, i.e., an rpm (the air flow rate) of the second blower27through the ECU25.

The gas phase refrigerant, which is outputted from the second evaporator18, is drawn into the suction inlet14cof the ejector14. The refrigerant flow of the arrow B, which flows in the refrigerant circulation passage11, is supplied to a refrigerant inlet (a drive flow inlet) of the nozzle portion14aof the ejector14. The refrigerant is depressurized and is expanded through the nozzle portion14a. Thus, the pressure energy of the refrigerant is converted into the velocity energy in the nozzle portion14aand is discharged from the outlet of the nozzle portion14at the high speed. Due to the decrease in the refrigerant pressure, the gas phase refrigerant, which is evaporated in the second evaporator18, is drawn through the suction inlet14c.

The refrigerant, which is discharged from the nozzle portion14a, and the refrigerant, which is drawn into the suction inlet14c, are mixed at the downstream side of the nozzle portion14aand are then supplied to the diffuser portion14b. Due to the increase in the cross sectional area of the refrigerant passage in the diffuser portion14b, the velocity energy (expansion energy) of the refrigerant is converted to the pressure energy. Thus, the pressure of the refrigerant is increased. The refrigerant discharged from the diffuser portion14bof the ejector14is supplied to the first evaporator15.

In the first evaporator15, the refrigerant absorbs heat from the conditioning air to be discharged into the vehicle passenger compartment, so that the refrigerant evaporates. After the evaporation, the gas phase refrigerant is drawn into the compressor12and is compressed. Thereafter, the refrigerant is discharged from the compressor12and flows in the direction of the arrow A in the refrigerant circulation passage11. Here, the ECU25controls the volume (displacement) of the compressor12to control the refrigerant discharge rate of the compressor12, so that the flow rate of the refrigerant supplied to the first evaporator15is adjusted. Furthermore, the ECU25controls the rpm (the air flow rate) of the first blower26to control the cooling capacity for cooling the subject cooling space, which is cooled by the first evaporator15, more specifically the cooling capacity for cooling the vehicle passenger compartment.

Next, advantages of the first embodiment will be described.

(1) The first evaporator15is arranged downstream of the diffuser portion14bof the ejector14, and the first branched passage16branches from the refrigerant circulation passage11at the downstream side of the radiator13and is connected to the suction inlet14cof the ejector14. The first flow rate control valve17and the second evaporator18are arranged in the first branched passage16. Therefore, the cooling operation can be simultaneously performed at both the first and second evaporators15,18.

(2) The refrigerant evaporation pressure of the first evaporator15is the pressure after the pressurization through the diffuser portion14b. In contrast, the outlet of the second evaporator18is connected to the suction inlet14cof the ejector14. Thus, the lowest pressure right after the depressurization at the nozzle portion14acan be applied to the outlet of the second evaporator18.

In this way, the refrigerant evaporation pressure (the refrigerant evaporation temperature) of the second evaporator18can be made lower than that of the first evaporator15. Thus, the cooling operation at the higher temperature range, which is suitable for cooling the vehicle passenger compartment, can be performed by the first evaporator15. At the same time, the cooling operation at the lower temperature range, which is lower than the higher temperature range and is suitable for cooling the interior of the refrigerator, can be performed by the second evaporator18.

As described above, even with the above simple structure, in which the first branched passage16is added, the cooling operation at the higher temperature range, which is suitable for cooling the vehicle passenger compartment, and the cooling operation at the lower temperature range, which is suitable for cooling the interior of the refrigerator, can be implemented. That is, the cooling operations at the two different temperature ranges can be implemented.

(3) As discussed above, the flow rate of the refrigerant supplied to the first evaporator15can be controlled by controlling the refrigerant discharge rate of the compressor12. Furthermore, the cooling capacity of the first evaporator15can be controlled by controlling the air flow rate of the first blower26.

Furthermore, the cooling capacity of the second evaporator18can be controlled by controlling the refrigerant flow rate through the first flow rate control valve17and by controlling the air flow rate of the second blower27.

As discussed above, the cooling capacity of the first evaporator15and the cooling capacity of the second evaporator18can be individually controlled. Thus, it is relatively easy to correspond to a change in the thermal load in the first and second evaporators15,18.

(4) The depressurized two-phase refrigerant, which is depressurized through the first flow rate control valve17and includes both the gas phase refrigerant and the liquid phase refrigerant, can be supplied to the second evaporator18through the first branched passage16. Thus, unlike Japanese Patent No. 3322263 ofFIG. 26, there is no need to provide the gas-liquid separator63at the downstream side of the first evaporator61to supply the liquid phase refrigerant to the second evaporator62.

Furthermore, as discussed above, the control of the refrigerant flow rate at the first evaporator15and the control of the refrigerant flow rate at the second evaporator18can be individually performed through the control of the refrigerant discharge rate of the compressor12and also through the control of the valve opening degree of the first flow rate control valve17. Thus, the control of the refrigerant flow rate of each evaporator15,18can be appropriately performed based on its thermal load. Therefore, the refrigerant flow rate can be adjusted in such a manner that the entire refrigerant becomes the gas phase refrigerant at the first evaporator15, which is located downstream of the ejector14.

Therefore, according to the present embodiment, it is possible to eliminate the gas-liquid separator63, which is required in Japanese Patent No. 3322263 ofFIG. 26. As a result, the manufacturing costs of the vapor compression cycle can be reduced.

(5) The pressure of the refrigerant is increased by the diffuser portion14bof the ejector14, so that the intake refrigerant pressure of the compressor12can be increased. In this way, the drive power for driving the compressor12can be minimized to improve the efficiency of the cycle.

SECOND EMBODIMENT

FIG. 2shows a vapor compression cycle of a second embodiment, which is similar to that of the first embodiment except first and second solenoid valves (a first opening and closing means and a second opening and closing means)19,20. The first solenoid valve19opens and closes the refrigerant circulation passage11on the upstream side of the ejector14. The second solenoid valve20opens and closes the first branched passage16on the upstream side of the first flow rate control valve17. Similar to the pressure control electromagnetic device12aof the compressor12, the opening and closing of the first and second solenoid valves19,20are controlled by a corresponding signal supplied from the ECU25.

Selection of an operating mode conducted by the ECU25will be described with reference toFIG. 3. First, user input information, temperature information of each subject cooling space and the temperature information of each evaporator15,18are inputted to the ECU25at step S110. The user information includes, for example, presence of a need (ON, OFF) for cooling the subject cooling space and the desired set temperature of the subject cooling space.

Next, at step S120, a target temperature of each subject cooling space or a target temperature of each evaporator15,18is determined by the ECU25based on the information inputted at step S110. In this way, the subject evaporator(s), which needs to be operated to achieve the required cooling capacity by supplying the refrigerant therethrough, is determined. Based on the respective target temperature, the best operating mode is determined with reference to, for example,FIG. 4at step S130.

In the present embodiment, a first evaporator operating mode (FIRST EVAPO. inFIG. 4), a second evaporator operating mode (SECOND EVAPO. inFIG. 4) and a multiple evaporator operating mode (MULTIPLE EVAPO. inFIG. 4) are provided. In the first evaporator operating mode, only the first evaporator15is operated to attain its cooling capacity. In the second evaporator operating mode, only the second evaporator18is operated to attain its cooling capacity. In the multiple evaporator operating mode, both the first and second evaporators15,18are operated to attain its cooling capacity.

For example, when a user actuates the cycle and sets the temperature of the subject cooling space, which is cooled by the first evaporator15, i.e., when the first evaporator15needs to be operated to attain its cooling capacity, the first evaporator operating mode is selected. Based on the selected operating mode, the ECU25controls the first and second solenoid valves19,20, the first flow rate control valve (FIRST CONTROL VALVE inFIG. 4)17and the first and second blowers26,27in the manner indicated inFIG. 4.

Thereafter, the ECU25controls the electrical device(s), such as the pressure control electromagnetic device12aof the compressor12, to adjust the temperature of the subject cooling space to the set temperature at step S140. In the above described manner, each operating mode shown inFIG. 4can be selected and can be set by the ECU25.

This point will be more specifically described. At the time of the first evaporator operating mode, the ECU25opens the first solenoid valve19and closes the second solenoid valve20. Then, the ECU25controls the volume (the refrigerant discharge rate) of the compressor12to control the flow rate of the refrigerant supplied to the first evaporator15. In this way, it is possible to adjust the total amount of heat, which is absorbed by the refrigerant at the first evaporator15from the air to be discharged into the subject cooling space. Furthermore, the flow rate of the cooling air into the subject cooling space, which is cooled by the first evaporator15, is controlled by controlling the rpm (the air flow rate) of the first blower26. In this way, the cooling capacity of the first evaporator15(more specifically, the cooling capacity for cooling the vehicle passenger compartment) is adjusted.

Furthermore, in the second evaporator operating mode, the ECU25closes the first solenoid valve19and opens the second solenoid valve20. The flow rate of the refrigerant, which is supplied to the second evaporator18, is controlled by controlling the volume (the refrigerant discharge rate) of the compressor12and the valve opening degree of the first flow rate control valve17.

Furthermore, the flow rate of the cooling air into the subject cooling space, which is cooled by the second evaporator18, is controlled by controlling the rpm (the air flow rate) of the second blower27. In this way, the cooling capacity of the second evaporator18(more specifically, the cooling capacity for cooling the interior of the refrigerator) is controlled.

Furthermore, in the multiple evaporator operating mode, the ECU25opens both the first and second solenoid valves19,20. Then, the ECU25controls the volume (the refrigerant discharge rate) of the compressor12to control the flow rate of the refrigerant supplied to the first evaporator15. The flow rate of the refrigerant, which is supplied to the second evaporator18, is adjusted by adjusting the valve opening degree of the first flow rate control valve17.

In addition, by individually controlling the rpm (the air flow rate) of the first blower26and the rpm (the air flow rate) of the second blower27, the flow rate of the cooling air discharged into the subject cooling space of the first evaporator15and the flow rate of the cooling air discharged into the subject cooling space of the second evaporator18are individually controlled. In this way, the cooling capacity of the first evaporator15and the cooling capacity of the second evaporator18are individually controlled.

When the pressure of the refrigerant, which is supplied to the ejector14, is increased by increasing the volume (the refrigerant discharge rate) of the compressor12, the suction capacity of the ejector14for drawing the gas phase refrigerant, which is evaporated in the second evaporator18, is increased. Even in this way, the flow rate of the refrigerant, which flows through the second evaporator18, can be controlled.

Furthermore, in the second evaporator operating mode, the refrigerant is supplied only to the second evaporator18, so that the refrigeration oil, which remains in the second evaporator18, can be returned to the compressor12.

THIRD EMBODIMENT

FIG. 5shows a vapor compression cycle according to a third embodiment. The vapor compression cycle of the third embodiment is similar to that of the second embodiment except a second branched passage23. The second branched passage23connects between a portion (a branching point) of the first branched passage16, which is on the upstream side of the first flow rate control valve17, and a portion (a merging point) of the refrigerant passage11, which connects between the first evaporator15and the compressor12.

A second flow rate control valve (a second metering means)24and a third solenoid valve (a third opening and closing means)28are arranged in the second branched passage23. The second flow rate control valve24controls the flow rate of the refrigerant and depressurizes the refrigerant. The third solenoid valve28opens and closes the second branched passage23. Furthermore, a third evaporator22is arranged on the downstream side of the second flow rate control valve24in the refrigerant flow direction in the second branched passage23. The air of a subject cooling space of the third evaporator22is blown by an electric blower (a third blower)29toward the third evaporator22.

Here, a downstream side of the third evaporator22is connected to a downstream side of the first evaporator15and is thus connected to the suction inlet side of the compressor12. Thus, the refrigerant evaporation pressure of the first evaporator15and the refrigerant evaporation pressure of the third evaporator22are generally the same as the suction pressure of the compressor12. Thus, the refrigerant evaporation temperature of the first evaporator15and the refrigerant evaporation temperature of the third evaporator22are also the same.

Therefore, for example, a front seat side space of the vehicle passenger compartment can be set as the subject cooling space of the first evaporator15, and a rear seat side space of the vehicle passenger compartment can be set as the subject cooling space of the third evaporator22. In this way, the front seat side space and the rear seat side space of the vehicle passenger compartment can be simultaneously cooled by the first and third evaporators15,22, respectively.

In the third embodiment, the second flow rate control valve24, the third solenoid valve28and the third blower29are also controlled by a corresponding control signal, which is supplied from the ECU25.

The control operation of the ECU25of the third embodiment is substantially the same as that of the second embodiment except step S130ofFIG. 3. That is, in the second embodiment, the operating mode is determined with reference toFIG. 4. In contrast, in the third embodiment, the operating mode is determined with reference toFIG. 6.

In the third embodiment, the number of controlled elements, which are controlled by the ECU25, is increased in comparison to that of the second embodiment, so that the number of the operating modes is increased, as shown inFIG. 6. However, similar to the second embodiment, the control flow of the ECU25is determined based on the operating mode of the corresponding required evaporator(s), which is required to attain the required cooling capacity (see S130inFIG. 3).

The operating modes of the third embodiment will be described further. The first evaporator operating mode (FIRST EVAPO.) and the second evaporator operating mode (SECOND EVAPO.) of the third embodiment are similar to those of the second embodiment. In the third evaporator operating mode (THIRD EVAPO.), the ECU25closes the first and second solenoid valves19,20and opens the third solenoid valve28.

The flow rate of the refrigerant, which is supplied to the third evaporator22, is controlled by controlling the volume (the refrigerant discharge rate) of the compressor12and the valve opening degree of the second flow rate control valve (SECOND CONTROL VALVE)24. Furthermore, the flow rate of the cooling air into the subject cooling space of the third evaporator22is controlled by controlling the rpm (the air flow rate) of the third blower29. In this way, the cooling capacity of the third evaporator22(more specifically, the cooling capacity for cooling the rear seat side space of the vehicle passenger compartment) is controlled.

In a first and second evaporator operating mode (FIRST, SECOND EVAPO. inFIG. 6), the ECU25opens the first and second solenoid valves19,20and closes the third solenoid valve28. The compressor12, the first flow rate control valve17and the first and second blowers26,27are controlled in a manner similar to that of the multiple evaporator operating mode of the second embodiment to control the cooling capacities of the first and second evaporators15,18.

In a first and third evaporator operating mode (FIRST, THIRD EVAPO. inFIG. 6), the ECU25opens the first and third solenoid valves19,28and closes the second solenoid valve20. Then, the flow rate of the refrigerant, which is supplied to the first evaporator15, is controlled by controlling the volume (the refrigerant discharge rate) of the compressor12. Also, the flow rate of the refrigerant, which is supplied to the third evaporator22, is controlled by controlling the valve opening degree of the second flow rate control valve24. Furthermore, the flow rate of the cooling air into the subject cooling space of the first evaporator15and the flow rate of the cooling air into the subject cooling space of the third evaporator22are controlled by controlling the rpm (the air flow rate) of the first blower26and the rpm (the air flow rate) of the third blower29, respectively. In this way, the cooling capacity of the first evaporator15and the cooling capacity of the third evaporator22are controlled.

In a second and third evaporator operating mode (SECOND, THIRD EVAPO.), the ECU25opens the second and third solenoid valves20,28and closes the first solenoid valve19. The cooling capacity of the second evaporator18and the cooling capacity of the third evaporator22are controlled by controlling the volume (the refrigerant discharge rate) of the compressor12, the valve opening degrees of the first and second flow rate control valves17,24and the air flow rates of the second and third blowers27,29.

In a first to third evaporator operating mode (ALL EVAPO. inFIG. 6), the ECU25opens all of the first to third solenoid valves19,20,28. Then, the flow rate of the refrigerant, which is supplied to the first evaporator15, is controlled by controlling the volume (the refrigerant discharge rate) of the compressor12. Also, the flow rate of the refrigerant to the second evaporator18and the flow rate of the refrigerant to the third evaporator22are controlled by controlling the valve opening degrees of the first and second flow rate control valve17,24, respectively.

Furthermore, the rpm's (the air flow rates) of the first to third blowers26,27,29are controlled to control the flow rates of the cooling air discharged into the corresponding subject cooling spaces, respectively. In this way, the cooling capacity of the first evaporator15, the cooling capacity of the second evaporator18and the cooling capacity of the third evaporator22are individually controlled.

In the above described manner, each operating mode shown inFIG. 6can be selected and can be set by the ECU25. Thus, the common subject cooling space or the multiple subject cooling spaces can be controlled by one or more of the three evaporators15,18,22.

Furthermore, in the second evaporator operating mode, the refrigerant is supplied only to the second evaporator18. Also, in the third evaporator operating mode, the refrigerant is supplied only to the third evaporator22. Thus, the refrigerant remained in the second evaporator18or the third evaporator22can be returned to the compressor12.

FOURTH EMBODIMENT

FIG. 7shows a vapor compression cycle of a fourth embodiment. The vapor compression cycle of the fourth embodiment is similar to the vapor compression cycle of the first embodiment except a third branched passage21. The third branched passage21extends from a portion (a branching point) of the refrigerant circulation passage11, which is located between the ejector14and the first evaporator15, to another portion (a merging point) of the refrigerant circulation passage11, which is located between the first evaporator15and the compressor12. A fourth evaporator (or a third evaporator)30is arranged in the third branched passage21. A fourth blower (or a third blower)31, which is an electric blower, is arranged to oppose the fourth evaporator30.

In this way, in addition to the first and second evaporators15,18, a predetermined subject cooling space can be cooled by the fourth evaporator30. Here, a downstream side of the fourth evaporator30is connected to a downstream side of the first evaporator15and is thus connected to the suction inlet side of the compressor12. Thus, the refrigerant evaporation pressure of the first evaporator15and the refrigerant evaporation pressure of the fourth evaporator30are generally the same as the suction pressure of the compressor12. Thus, the refrigerant evaporation temperature of the first evaporator15and the refrigerant evaporation temperature of the fourth evaporator30are also the same.

Even in the fourth embodiment, similar to the third embodiment, the common subject cooling space or the multiple subject cooling spaces can be cooled by the three evaporators15,18,30.

FIFTH EMBODIMENT

In each of the first to fourth embodiments, the ejector14and the first evaporator15are connected in series. Thus, the ejector14has the flow rate adjusting function for adjusting the flow rate of the refrigerant to the first evaporator15and also has the pumping function for creating a refrigerant pressure difference between the first evaporator15and the second evaporator18.

Therefore, at the time of designing the ejector14, the required specification for achieving both the flow rate adjusting function and the pumping function should be satisfied. Thus, in order to achieve the flow rate adjusting function for adjusting the flow rate of the refrigerant to the first evaporator15, the design needs to rely on the first evaporator15. As a result, the operation of the vapor compression cycle at the high efficiency becomes difficult.

Thus, in the fifth embodiment, the ejector14has only the pumping function without the flow rate adjusting function for adjusting the flow rate of the first evaporator15to allow easy designing of the ejector14, which enables the highly efficient operation of the vapor compression cycle.

The fifth embodiment will be described more specifically with reference toFIG. 8. In the refrigerant circulation passage11, a dedicated metering mechanism (a first metering means)32is provided between the outlet of the radiator13and the inlet of the first evaporator15. Furthermore, in the fifth embodiment, the ejector14is not provided in the refrigerant circulation passage11. Rather, the ejector14is provided in parallel with the metering mechanism32.

Although various devices can be used as the metering mechanism32, a thermostatic expansion valve, which controls its valve opening degree in a manner that keeps the superheat of the refrigerant at the outlet of the first evaporator15at a predetermined value, is used as the metering mechanism32in the present embodiment.

A metering mechanism (a second metering means)17and the second evaporator18are arranged in series in the first branched passage16, which branches from the portion of the refrigerant circulation passage11between the outlet of the radiator13and the inlet of the ejector14. Furthermore, the outlet of the second evaporator18is connected to the suction inlet14cof the ejector14. Although various devices can be used as the metering mechanism17of the first branched passage16, a fixed metering device, such as a capillary tube of a simple structure, is used as the metering mechanism17in this embodiment.

Next, operation of the fifth embodiment will be described. When the compressor12is operated, the discharged refrigerant, which is discharged from the compressor12, releases the heat to the external air and is condensed in the radiator13. Thereafter, the flow of the condensed refrigerant is divided into the following three flows.

That is, the first refrigerant flow passes the metering mechanism32and is depressurized. Then, the first refrigerant flow enters the first evaporator15. The second refrigerant flow passes the nozzle portion14aof the ejector14and is depressurized. Then, the second refrigerant flow passes the diffuser portion14band is pressurized. Thereafter, the second refrigerant flow enters the first evaporator15. The third refrigerant flow passes the metering mechanism17and is depressurized. Thereafter, the third refrigerant flow passes the second evaporator18and is then drawn into the suction inlet14cof the ejector14.

Even in the fifth embodiment, the ejector14performs the pumping function. That is, the ejector14draws the refrigerant present at the outlet of the second evaporator18and mixes the drawn refrigerant with the refrigerant flow (drive flow), which has passed the nozzle portion14a, so that the mixed refrigerant is pressurized at the diffuser portion14b. Thus, the evaporation pressure of the first evaporator15is higher than the evaporation pressure of the second evaporator18, so that the pressure difference (the refrigerant evaporation temperature difference) is created between the evaporation pressure of the second evaporator18and the evaporation pressure of the first evaporator15.

The flow rate of the refrigerant, which enters the first evaporator15, can be controlled through the dedicated metering mechanism32. Thus, the ejector14does not need to have the flow rate adjusting function for adjusting the flow rate of the first evaporator15. Similarly, the flow rate of the refrigerant, which enters the second evaporator18, is controlled through the dedicated metering mechanism17. Thus, the function of the ejector14is specialized to the pumping function for creating the pressure difference between the first evaporator15and the second evaporator18.

In this way, the configuration of the ejector14can be designed to create the predetermined pressure difference between the first evaporator15and the second evaporator18, i.e., to set the flow rate of the refrigerant in the ejector14to the predetermined flow rate. As a result, the vapor compression cycle can be operated at the high efficiency even when the cycle operational condition (e.g., the rpm of the compressor, the external air temperature, the subject cooling space temperature) varies through a wide range.

Furthermore, the function of the ejector14is specialized only to the pumping function, so that it is relatively easy to use the fixed nozzle, which has the fixed passage cross sectional area, as the nozzle portion14aof the ejector14. The use of the fixed nozzle allows a reduction in the manufacturing costs of the ejector14.

SIXTH EMBODIMENT

FIG. 9shows a sixth embodiment, which is a modification of the fifth embodiment. Specifically, in the sixth embodiment, as shown inFIG. 9, the downstream side (the outlet) of the ejector14is connected to the downstream side (the outlet) of the first evaporator15. Even with this modification, the vapor compression cycle can be operated at the high efficiency due to the appropriate design of the configuration of the ejector14.

However, in the sixth embodiment, the refrigerant flow (the drive flow), which has passed the nozzle portion14aof the ejector14, is directly drawn into the compressor12without passing through any evaporator, so that a problem of liquid refrigerant return to the compressor12(sometimes referred to as “liquid slugging” of the compressor) possibly occurs.

Therefore, it is preferred to apply the sixth embodiment to the case where the flow rate of the drive flow in the ejector14is relatively small, i.e., to the case where the capacity of the second evaporator18is small.

In the sixth embodiment, when a thermostatic expansion valve, which controls its valve opening degree in a manner that keeps the superheat of the refrigerant at the downstream side of the ejector14at the predetermined value, is used, the liquid refrigerant return from the portion of the refrigerant passage, which is located on the downstream side of the ejector14, to the compressor12can be more reliably limited.

SEVENTH EMBODIMENT

FIG. 10shows a seventh embodiment, which is a modification of the sixth embodiment. Specifically, in the seventh embodiment, with reference toFIG. 10, the ejector14, the metering mechanism17and the second evaporator18, which are located within a dotted line frame in the drawing, are pre-assembled as an integral unit33.

Two pipe lines, which respectively constitute an inlet passage portion of the first branched passage16and a downstream side passage portion located downstream of the ejector14, are provided to the integral unit33. In this way, the known vapor compression refrigeration cycle, which has the refrigerant circulation passage11(including the compressor12, the radiator13, the metering mechanism32and the first evaporator15), can be easily modified to the vapor compression cycle, which includes the two evaporators15,18.

Although the seventh embodiment is the modification of the sixth embodiment, the aspect of the integral unit33of the seventh embodiment may be implemented in the fifth embodiment (FIG. 8).

EIGHTH TO TENTH EMBODIMENTS

In eighth to tenth embodiments, the aspect of the fifth embodiment (FIG. 8) is implemented in the vapor compression cycle, which has the three evaporators15,18,22.

FIG. 11shows the eighth embodiment, in which the aspect of the fifth embodiment (FIG. 8) is applied to the third embodiment shown inFIG. 5.

FIG. 12shows the ninth embodiment, in which the downstream side passage located downstream of the ejector14is connected between the downstream side of a metering mechanism (a third metering means)24and the upstream side of the third evaporator22in the eighth embodiment shown inFIG. 11.

FIG. 13shows the tenth embodiment, in which the downstream side passage located downstream of the ejector14is directly connected to the suction inlet of the compressor12in the eight embodiment shown inFIG. 11. The above point is similar to that of the sixth and seventh embodiments shown inFIGS. 9 and 10.

Even in the eighth to tenth embodiments, the refrigerant evaporation pressure (the refrigerant evaporation temperature) of the first evaporator15becomes the same as that of the third evaporator22, and the refrigerant evaporation pressure (the refrigerant evaporation temperature) of the second evaporator18becomes smaller than that of the first and the third evaporators15,22.

Furthermore, in the eighth to tenth embodiments, the function of the ejector14can be specialized to the pumping function, so that the vapor compression cycle can be operated at the high efficiency upon the appropriate designing of the configuration of the ejector14.

In any of the first to tenth embodiments, the basic cycle structure is the same as that of the first embodiment, so that the advantages similar to those recited in (1) to (5) in the first embodiment can be achieved.

ELEVENTH EMBODIMENT

An eleventh embodiment will be described with reference toFIGS. 14 to 17B.FIG. 14schematically shows a vapor compression cycle, in which a refrigeration cycle device according to the eleventh embodiment of the present invention is implemented and which is suitable for a refrigeration cycle of a vehicle air conditioning system. In the vapor compression cycle, a refrigerant circulation passage R is provided. A compressor101for drawing and compressing refrigerant is arranged in the refrigerant circulation passage R. In the refrigerant circulation passage R, a radiator (a high pressure side heat exchanger)102is arranged downstream of the compressor101. The radiator102releases the heat of the high pressure refrigerant, which is discharged from the compressor101.

The refrigerant, which is discharged from the radiator102, is supplied to a first refrigerant passage111of the refrigeration cycle device of the present embodiment. The refrigeration cycle device of the present embodiment includes a box type thermostatic expansion valve105and an ejector103. More specifically, a refrigerant inlet103aof the ejector103(i.e., a refrigerant inlet103aof a nozzle portion131of the ejector103) is air-tightly connected to a downstream side of a metering portion S1of the expansion valve105, i.e., to an outlet of the first refrigerant passage111. Since the expansion valve105and the ejector103are main features of the present embodiment, structures of the expansion valve105and of the ejector103will be described in greater detail.

In the refrigeration cycle device, a first evaporator104is connected to the refrigerant discharge outlet103cof the ejector103on the downstream side of the ejector103. In the first evaporator104, the refrigerant, which is discharged from the refrigerant discharge outlet103c, is evaporated. A refrigerant outlet of the first evaporator104is connected to a suction inlet of the compressor101through a second refrigerant passage112of the refrigeration cycle device. The flow of the refrigerant is divided into two flows at a location (a branching point) between the radiator102and the refrigeration cycle device (i.e., the expansion valve105and the ejector103). One of the two divided flows is conducted through a refrigerant circulation passage R1and is supplied to an inlet of the first refrigerant passage111of the refrigeration cycle device. The other one of the two divided flows is conducted through a branched passage R2and is supplied to a refrigerant suction inlet103bof the refrigeration cycle device (more specifically, the ejector103).

Next, the details of the structures of the expansion valve105and of the ejector103will be described.FIG. 15is a cross sectional view of the expansion valve105of the present embodiment. The expansion valve105is arranged in the refrigerant passage between the radiator102and the ejector103, i.e., is arranged on the upstream side of a nozzle portion131of the ejector103. The expansion valve105depressurizes and expands the high pressure refrigerant, which is discharged from the radiator102, to two-phase refrigerant of a gas and liquid mixture. The expansion valve105of the present embodiment has the structure similar to that of a know box type thermostatic expansion valve. A valve opening degree of the expansion valve105is controlled to keep the refrigerant superheat in a predetermined range (e.g., 0.1 degrees to 10 degrees) at the refrigerant outlet of the first evaporator104.

The expansion valve105includes a valve block (a valve main body) D, an element arrangement E, a heat conducting portion120, a conducting rod125and a ball valve element110. The valve block D is made of, for example, aluminum and is formed into a generally rectangular parallelepiped body. Furthermore, the valve block D includes the first refrigerant passage111and the second refrigerant passage112.

The first refrigerant passage111includes an inflow port (refrigerant inlet)111a, an outflow port (refrigerant outlet)111band a communication hole111c. The inflow port111ais connected to the outlet of the radiator102. The outflow port111bis connected to a refrigerant inlet103aof the ejector103. The communication hole111ccommunicates between the inflow port111aand the outflow port111b. A conical valve seat surface111dis provided to an inlet of the communication hole111con the inflow port111aside of the communication hole111c. The second refrigerant passage112includes an inflow port (refrigerant inlet)112a, an outflow port (refrigerant outlet)112band a communication passage112c. The inflow port112ais connected to the outlet of the evaporator104. The outflow port112bis connected to the suction inlet of the compressor101. The communication passage112ccommunicates between the inflow port112aand the outflow port112band also communicates with the heat conducting portion120.

The element arrangement E includes a diaphragm113, a receiving portion114and a cover portion115. The diaphragm113is made of a flexible thin metal plate. The receiving portion114holds the diaphragm113. The element arrangement E is screwed and secured to a top of the valve block D through a packing116. The receiving portion114and the cover portion115are connected together by, for example, TIG welding. The diaphragm113and the cover portion115form a diaphragm chamber117.

Saturated gas, which is of the same type as the refrigerant gas is used in the refrigeration cycle, is filled in the diaphragm chamber117. A through hole for filling the saturated gas into the diaphragm chamber117penetrates through the cover portion115. After filling of the saturated gas into the diaphragm chamber117, a plug118is fitted to the through hole of the cover portion115to air-tightly close it. Each component (the diaphragm113, the receiving portion114, the cover portion115and the plug118) of the element arrangement E is made of a common metal material (e.g. stainless steel), which serves as a first material.

The heat conducting portion120is made of a metal material (e.g., aluminum or brass), which serves as a second material and shows a relatively high thermal conductivity that is higher than that of the first material, and is formed into a cylindrical body. An upper surface of the cylindrical body of the heat conducting portion120is urged upwardly by an urging force (described later) and is closely engaged with a lower surface of the diaphragm113. A change in the temperature of the refrigerant (gas phase refrigerant evaporated in the evaporator104), which flows in the second refrigerant passage112, is conducted to the diaphragm113through the heat conducting portion120. Furthermore, a lower surface of the cylindrical body of the heat conducting portion120is engaged with the conducting rod125to conduct displacement of the diaphragm113to the ball valve element110in cooperation with the conducting rod125.

The conducting rod125is arranged under the heat conducting portion120and is slidably held by the valve block D. The conducting rod125is engaged with the lower surface of the heat conducting portion120at its top end. Furthermore, the conducting rod125extends through the second refrigerant passage112(the communication passage112c) in the vertical direction and is inserted in to the communication hole111cof the first refrigerant passage111. A lower end of the conducting rod125is engaged with a top surface of the ball valve element110, which is urged against the conical seat surface111dby a spring122. In a portion of the valve block D between the first refrigerant passage111and the second refrigerant passage112, an O-ring (a seal portion)119is provided to the conducting rod125, which is vertically slidably received in the valve block D.

As shown inFIG. 15, the ball valve element110is provided to the inlet of the communication hole111cand is held between the conducting rod125and a valve receiving member121. When the ball valve element110is seated against the seat surface111d, the ball valve element110closes the communication hole111c. When the ball valve element110is lifted away from the seat surface111d, the ball valve element110opens the communication hole111c. InFIG. 15, the ball valve element110is stationary held in the position where the urging force for downwardly urging the diaphragm113(the pressure of the diaphragm chamber117—the pressure of the refrigerant vapor applied to the lower side of the diaphragm113) and the load of the spring122, which urges the ball valve element110in the upward direction inFIG. 15through the valve receiving member121, are balanced.

The spring122is arranged between the valve receiving member121and an adjusting screw123, which is installed to the lower end of the valve block D. The spring122urges the ball valve element110in the upward direction (the direction for reducing the valve opening degree) inFIG. 15through the valve receiving member121. The adjusting screw123adjusts the valve opening pressure of the ball valve element110(the load of the spring122that urges the ball valve element110) and is threadably engaged with the lower end of the valve block D through an O-ring124.

Next, operation of the expansion valve105will be described. The flow rate of the refrigerant, which passes the communication hole111c, is determined based on the valve opening degree of the ball valve element110, i.e., based on the position (the lift amount) of the ball valve element110relative to the seat surface111d. The ball valve element110is moved to a balanced position where the pressure of the diaphragm chamber117, which urges the diaphragm113in the downward direction inFIG. 15, the load of the spring122, which urges the diaphragm113in the upward direction inFIG. 15, and the low pressure in the cycle (the pressure of the refrigerant vapor applied to the lower side of the diaphragm113) are balanced.

When the temperature of the vehicle passenger compartment is increased from the stable state where the vapor pressure is stable, and thereby the refrigerant is rapidly evaporated in the evaporator104, the temperature (the superheat) of the refrigerant vapor at the outlet of the evaporator104is increased. In this way, the change in the temperature of the refrigerant vapor, which flows in the second refrigerant passage112, is conducted to the sealed gas, which is sealed in the diaphragm chamber117, through the heat conducting portion120and the diaphragm113. When the temperature of the sealed gas in the diaphragm chamber117is increased, the pressure of the diaphragm chamber117is increased.

Thus, the diaphragm113is urged and is moved in the downward direction inFIG. 15. As a result, the valve opening degree is increased, and the flow rate of the refrigerant supplied to the evaporator104is increased. In contrast, when the temperature of the passenger compartment is decreased, and the superheat of the outlet of the evaporator104is decreased, the change in the temperature of the refrigerant vapor, which flows in the second refrigerant passage112, is conducted to the sealed gas of the diaphragm chamber117. Due to the decrease in the temperature of the sealed gas, the pressure of the diaphragm chamber117is decreased.

As a result, when the diaphragm113is pushed in the upward direction inFIG. 15, and thereby the ball valve element110is moved in the upward direction inFIG. 15, the valve opening degree is decreased. Therefore, the flow rate of the refrigerant, which is supplied to the evaporator104, is decreased. Therefore, during the normal cycle operation, the valve opening degree is controlled to make the temperature (the superheat) of the refrigerant vapor, for example, about 5 degrees Celsius and thereby to control the flow rate of the refrigerant, which flows in the communication hole111c.

FIG. 16is a cross sectional view of the structure of the ejector103of the present embodiment, andFIGS. 17A and 17Bare descriptive views for describing the advantages of the ejector103ofFIG. 16. The ejector103depressurizes and expands the refrigerant, which is supplied from the radiator102through the refrigerant inlet103avia the first refrigerant passage111(the first metering portion S1inFIG. 14) of the expansion valve105, so that the ejector103draws the gas phase refrigerant, which is evaporated in the second evaporator106, through the refrigerant suction inlet103b. Furthermore, the ejector103converts the expansion energy of the refrigerant to the pressure energy of the refrigerant and discharges the refrigerant from the refrigerant discharge outlet103cto increase the suction pressure of the compressor101.

The ejector103includes the nozzle portion131, the mixing portion132and the diffuser portion133. The nozzle portion131isentropically depressurizes and expands the high pressure refrigerant, which is supplied through the refrigerant inlet103a, by converting the pressure energy of the high pressure refrigerant, which is supplied through the refrigerant inlet103a, into the velocity energy. Through use of the entraining action of the high velocity refrigerant flow (drive flow), which is discharged from the nozzle portion131, the mixing portion132draws the gas phase refrigerant, which is evaporated in the second evaporator106, through the suction inlet103b. Then, the mixing portion132mixes the drawn refrigerant, which is drawn from the second evaporator106, and the discharged refrigerant, which is discharged from the nozzle portion131. The diffuser portion133further mixes the drawn refrigerant, which is drawn from the second evaporator106, and the discharged refrigerant, which is discharged from the nozzle portion131. Also, at the same time, the diffuser portion133converts the velocity energy of the refrigerant to the pressure energy of the refrigerant to increase the pressure of the refrigerant.

At this time, in the mixing portion132, the drive flow and the drawn flow are mixed such that the sum of the kinetic energy of the drive flow and kinetic energy of the drawn flow is conserved. Thus, even in the mixing portion132, the pressure (the static pressure) of the refrigerant is increased. In the diffuser portion133, the passage cross sectional area is progressively increased to convert the velocity energy (dynamic pressure) of the refrigerant to the pressure energy (static pressure). Thus, in the ejector103, the refrigerant pressure is increased in both of the mixing portion132and the diffuser portion133. Hereinafter, the mixing portion132and the diffuser portion133will be collectively referred to as a pressurizing portion.

In the present embodiment, a Laval nozzle, which has a throat (a second metering portion) S2that has the smallest passage cross sectional area in the Laval nozzle, is used to accelerate the velocity of the refrigerant, which is discharged from the nozzle portion131, to a sonic velocity or higher velocity. However, it should be understood that a tapered nozzle can be used in place of the Laval nozzle. In the present embodiment, the passage cross sectional area of the mixing portion132before the diffuser portion133is constant. Alternatively, the passage cross sectional area of the mixing portion132can be tapered to have an increasing passage cross sectional area, which increases toward the diffuser portion133.

The high pressure refrigerant, which is cooled in the radiator102, is isentropically depressurized to the two-phase refrigerant (mixture of gas and liquid) range. Thereafter, the refrigerant is isentropically depressurized and is expanded by the nozzle portion131of the ejector103and is supplied to the mixing portion132at the sonic velocity or higher velocity. Therefore, the refrigerant is boiled once in the expansion valve105and is expanded at the inlet of the nozzle portion131to recover the pressure. In this way, the refrigerant can be boiled in the nozzle portion131while the boiling nucleus is kept generated. Thus, the boiling of the refrigerant in the nozzle portion131is promoted, and the liquid refrigerant droplets are atomized to improve the ejector efficiency ηe (FIG. 17A).

In the present embodiment, chlorofluorocarbon is used as the refrigerant to keep the high pressure side refrigerant pressure (i.e., the pressure of the refrigerant supplied to the nozzle portion131) equal to or less than the critical pressure of the refrigerant. Due to the pump action, which utilizes the entraining action of the high velocity refrigerant that is supplied to the mixing portion132, the refrigerant, which is evaporated in the second evaporator106, is drawn into the mixing portion132. Thus, the low pressure side refrigerant is circulated through the second evaporator106and the pressurizing portion132,133of the ejector103in this order.

In contrast, the refrigerant (the drawn flow), which is drawn from the second evaporator106, and the refrigerant (the drive flow), which is discharged from the nozzle portion131, are mixed in the mixing portion132, and the dynamic pressure of the mixed refrigerant is converted into the static pressure in the diffuser portion133. Thereafter, the mixed refrigerant is discharged from the diffuser portion133. Therefore, in the present embodiment, the nozzle efficiency and the ejector efficiency are increased while achieving the sufficient refrigeration capacity, and it is possible to correspond to a wide range of the load change.

In the first evaporator104, heat is exchanged between the refrigerant and the air to be discharged into the passenger compartment, so that the refrigerant is evaporated upon absorbing the heat. In this way, the cooling capacity is implemented. Furthermore, in the second evaporator106, heat is exchanged between the refrigerant and the air in the interior of the refrigerator, so that the refrigerant is evaporated upon absorbing the heat. In this way, the cooling capacity is implemented.

Next, operation of the present embodiment will be described with reference to the above structure. When the compressor101is operated, the refrigerant is compressed in the compressor101, so that the high temperature and high pressure refrigerant is discharged from the compressor101and is supplied to the radiator102. In the radiator102, the high temperature refrigerant releases the heat to the external air, which is external to the vehicle passenger compartment. That is, in the radiator102, the refrigerant is cooled by the external air and is condensed to the liquid state.

The liquid phase refrigerant, which is outputted from the radiator102, is divided into the refrigerant circulation passage R1and the branched passage R2. In the refrigerant circulation passage R1, the refrigerant is supplied from the first refrigerant passage111of the refrigeration cycle device to the ejector103and is depressurized in the nozzle portion131. That is, in the nozzle portion131, the pressure energy of the refrigerant is converted into the velocity energy. The refrigerant, which is discharged from the outlet of the nozzle portion131at the high velocity, draws the gas phase refrigerant, which is evaporated in the second evaporator106, through the suction inlet103bdue to the adiabatic heat drop that occurs at the time of discharging the refrigerant from the nozzle portion131.

The discharged refrigerant, which is discharged from the nozzle portion131, and the drawn refrigerant, which is drawn from the second evaporator106, are mixed, and the mixed refrigerant is supplied to the diffuser portion133. At this time, the expansion energy of the refrigerant is converted into the pressure energy, so that the pressure of the refrigerant is increased. The refrigerant, which is discharged from the ejector103, is supplied to the first evaporator104. In the first evaporator104, the refrigerant absorbs the heat from the air to be discharged into the vehicle passenger compartment. In other words, in the first evaporator104, the refrigerant is heated by the interior air of the vehicle passenger compartment and is evaporated.

The evaporated gas phase refrigerant is supplied to the compressor101through the second refrigerant passage112of the refrigeration cycle device. In the branched passage R2, the other divided refrigerant flow is supplied to the second evaporator106. In the second evaporator106, the refrigerant absorbs the heat from the interior air of the refrigerator. In other words, in the second evaporator106, the refrigerant is heated by the interior air of the refrigerator and is evaporated. The evaporated refrigerant is drawn through the suction inlet103bof the ejector103.

Next, the characteristic features and advantages of the present embodiment will be described. The refrigeration cycle device of the present embodiment includes the box type thermostatic expansion valve105and the ejector103. The expansion valve105forms the first metering portion S1to serve as the depressurizing means for depressurizing the high pressure refrigerant. Furthermore, the expansion valve105adjusts the flow rate of the refrigerant, which passes the first refrigerant passage111, based on the superheat of the refrigerant, which passes the first refrigerant passage111. The ejector103includes the nozzle portion131and the pressurizing portion132,133. The nozzle portion131forms the second metering portion S2and depressurizes and expands the refrigerant by converting the pressure energy of the high pressure refrigerant, which is supplied through the inlet103a, into the velocity energy. The pressurizing portion132,133draws the gas phase refrigerant from the suction inlet103bthrough use of the high velocity refrigerant, which is discharged from the nozzle portion131. The pressuring portion132,133converts the velocity energy to the pressure energy while mixing the discharged refrigerant, which is discharged from the nozzle portion131, and the drawn refrigerant, which is drawn from the suction inlet103b, so that the pressure of the mixed refrigerant is increased by the pressurizing portion132,133. The refrigerant inlet103aof the ejector103is air-tightly connected to the downstream side of the metering portion S1of the box type thermostatic expansion valve105.

FIG. 14is the schematic view of the vapor compression cycle, which includes the refrigeration cycle device of the eleventh embodiment. With respect to the previously proposed refrigeration cycle, in the vapor compression cycle of the present embodiment, the ejector103, which includes the nozzle portion131and the pressurizing portion132,133, is placed between the expansion valve105and the first evaporator104and is connected to the expansion valve105, so that the ejector103draws and pressurizes the refrigerant, which is supplied from the second evaporator106. Therefore, the first and second evaporators104,106are operated at different temperature ranges. At this time, the ejector103is easily and detachably connected to the expansion valve105, so that the variable ejector having the simple structure is provided.

Furthermore, in order to correspond to the load change, the superheat at the outlet of the first evaporator104is sensed. At the time of the high load operation, the superheat becomes excessively large, and thereby the expansion valve105is opened. In contrast, at the time of the low load operation, the expansion valve105is closed. Therefore, the flow rate of the refrigerant is adjusted. Furthermore, the nozzle portion131converts the pressure energy to the velocity energy. However, when the two-phase refrigerant of the mixture of the gas and liquid is used, the nozzle efficiency is decreased due to the delay in the boiling of the refrigerant in the second metering portion S2. To address this issue, the boiling nucleus is initially generated in the expansion valve105by the depressurization to improve the ejector efficiency (the nozzle efficiency).

Furthermore, a predetermined space is interposed between the first metering portion S1and the second metering portion S2. In the case where the nozzle efficiency is improved by initially generating the boiling nucleus in the expansion valve105, the space between the first metering portion S1, which is implemented by the expansion valve105, and the second metering portion S2, which is implemented by the nozzle throat, contributes to the improved performance. In this way, through the simple assembly of the expansion valve105with the ejector103, and the provision of the predetermined space between the first metering portion S1and the second metering portion S2, the high ejector efficiency can be achieved.

Furthermore, the expansion valve105and the ejector103are connected to each other such that the center axis of the expansion valve105is perpendicular to the center axis of the ejector103. In this way, the direction of the suction inlet103bof the ejector103can be freely selected within360degrees to provide more freedom at the time of mounting the ejector103.

Furthermore, the box type expansion valve105includes the first refrigerant passage111, the second refrigerant passage112, the ball valve element110, the element arrangement E and the heat conducting portion120. The first refrigerant passage111is connected to the inlet of the first evaporator104. The second refrigerant passage112is connected to the outlet of the first evaporator104. The ball valve element110changes the flow rate of the refrigerant in the first refrigerant passage111. In the element arrangement E, the diaphragm113is held, i.e., is clamped between the receiving portion114and the cover portion115, and the saturated gas is sealed in the diaphragm chamber117between the diaphragm113and the cover portion115. Furthermore, the diaphragm113, the receiving portion114and the cover portion115are made of the common material. The element arrangement E is detachably connected to the valve block D. The heat conducting portion120is made of the different material, which shows the thermal conductivity that is higher than that of the element arrangement E. The heat conducting portion120conducts the temperature change of the refrigerant, which flows in the second refrigerant passage112, to the diaphragm113. Furthermore, the heat conducting portion120conducts the displacement of the diaphragm113to the ball valve element110. The flow rate of the refrigerant, which flows in the first refrigerant passage111, is adjusted based on the displacement of the ball valve element10.

The above structure is of the previously proposed box type expansion valve105. With the above structure, through the combination of the previously proposed devices, the manufacturing costs can be minimized. Furthermore, variations of the above structure can be implemented at the relatively low costs by appropriately combining the previously proposed devices.

Furthermore, the above vapor compression type refrigeration cycle, which transfers the heat of the lower temperature side to the higher temperature side, includes the compressor101, the radiator102, the refrigeration cycle device103,105, the first evaporator104, the branched passage R2and the second evaporator106. The compressor101compresses the refrigerant. The radiator102releases the heat from the high pressure refrigerant, which is discharged from the compressor101. The refrigeration cycle device supplies the refrigerant, which is outputted from the radiator102, to the first refrigerant passage111. The outlet of the first evaporator104is connected to the suction inlet of the compressor101through the second refrigerant passage112. The first evaporator104evaporates the refrigerant, which is discharged from the outlet103cof the refrigeration cycle device. The branched passage R2branches the refrigerant flow at the point (branching point) between the radiator102and the refrigeration cycle device and conducts it to the suction inlet103b. The second evaporator106is arranged in the branched passage R2and evaporates the refrigerant.

With the above structure, the ejector103is easily detachable relative to the expansion valve105. Thus, the vapor compression cycle can be made with the simple structure. Furthermore, in the case where the second evaporator106is not used, the simple normal expansion valve cycle can be made by simply removing the ejector103and the second evaporator106.

Furthermore, in the above embodiment, the refrigerant can be one of the chlorofluorocarbon refrigerant, hydrocarbon (HC) refrigerant and the carbon dioxide refrigerant. The chlorofluorocarbon is a general name of an organic compound, which is made from carbon, fluorine, chlorine and hydrogen. The chlorofluorocarbon has been widely used as the refrigerant. The chlorofluorocarbon refrigerant includes hydrochlorofluorocarbon (HCFC) refrigerant, hydrofluorocarbon (HFC) refrigerant and the like and is called as the alternative for chlorofluorocarbon, which is used to limit damage of the ozone shield.

The HC refrigerant is the natural refrigerant material, which includes hydrogen and carbon. Examples of the HC refrigerant include R600a (isobutene) and R290 (propane). Accordingly, any one of the chlorofluorocarbon refrigerant, the hydrocarbon refrigerant and the carbon dioxide refrigerant can be used as the refrigerant of the present embodiment.

TWELFTH EMBODIMENT

FIG. 18Ais a partial cross sectional view of a refrigeration cycle device according to a twelfth embodiment of the present invention, andFIG. 18Bis a view seen in a direction of XVIIIB inFIG. 18A. In the eleventh embodiment, the expansion valve105and the ejector103are connected to each other such that the center axis of the expansion valve105is perpendicular to the center axis of the ejector103. In the twelfth embodiment, the expansion valve105and the ejector103are connected to each other such that the center axis of the expansion valve105is parallel to the center axis of the ejector103. In this way, the direction of the refrigerant discharge outlet103cof the ejector103can be freely selected within 360 degrees to provide more freedom at the time of mounting the ejector103.

The first to twelfth embodiments can be modified as follows.

(1) In the first embodiment, the present invention is implemented in the vehicle air conditioning and refrigerating system. Alternatively, both the first evaporator15, which has the higher refrigerant evaporation temperature, and the second evaporator18, which has the lower refrigerant evaporation temperature, can be used to cool different regions (e.g., the vehicle front seat side region and the vehicle rear seat side region) of the vehicle passenger compartment.

(2) In the first embodiment, both the first evaporator15, which has the higher refrigerant evaporation temperature, and the second evaporator18, which has the lower refrigerant temperature, can be used to cool the interior of the refrigerator. More specifically, the first evaporator15, which has the higher refrigerant evaporation temperature, can be used to cool the interior of the chillroom of the refrigerator, and the second evaporator18, which has the lower refrigerant evaporation temperature, can be used to cool the interior of the freezing room of the refrigerator.

(3) The vapor compression cycle of the present invention can be applied to a vapor compression cycle, such as a heat pump of a water heater.

(4) Although the type of the refrigerant is no specified in the first to tenth embodiments, the refrigerant can be any suitable refrigerant, such as, chlorofluorocarbon, hydrocarbon (HC) alternatives for chlorofluorocarbon, carbon dioxide, which is applicable to both a supercritical vapor compression cycle and a sub-critical vapor compression cycle.

(5) In the first embodiment, the gas-liquid separator is not used. Alternatively, the gas-liquid separator may be provided on the upstream side of the first evaporator15to provide only the liquid phase refrigerant to the first evaporator15. Further alternatively, the gas-liquid separator may be provided to the upstream side of the compressor12to provide only the gas phase refrigerant to the compressor12. Furthermore, a receiver may be provided on the downstream side of the radiator13. The receiver separates the liquid phase refrigerant from the gas phase refrigerant and supplies only the liquid phase refrigerant to its downstream side.

(6) In the first to fourth embodiments, the first flow rate control valve17is provided on the upstream side of the second evaporator18. In a case where a change in the thermal load of the second evaporator18is relatively small, a fixed metering device, such as a capillary tube, which has an aperture of a fixed size, can be used as the first flow rate control valve17.

Furthermore, when the fixed metering device and the solenoid valve are integrated together as the first flow rate control valve17, it is possible to provide a metering mechanism, in which the flow rate control function of the fixed metering device and the flow passage closing (shutting off) function are combined.

Furthermore, the first flow rate control valve17can be a metering device (e.g., an expansion valve), which has a mechanism to control an opening degree of its aperture based on the sensed superheat at the outlet of the evaporator.

Furthermore, in the second and third embodiments, the first flow rate control valve17is separated from the second solenoid valve20, and the second flow rate control valve24is separated from the third solenoid valve28. Alternatively, in place of the combination of the first flow rate control valve17and the second solenoid valve20and/or the combination of the second flow rate control valve24and the third solenoid valve28, a metering valve(s) with the flow passage closing (shutting off) function, in which the flow rate control valve and the solenoid valve are integrated, can be used.

(7) In the first to fourth embodiments, the variable displacement compressor is used as the compressor12, and the volume of the variable displacement compressor12is controlled by the ECU25to control the refrigerant discharge rate. Alternatively, a fixed displacement compressor may be used as the compressor12. In such a case, on- and off-operations of the fixed displacement compressor12are controlled by an electromagnetic clutch, and a ratio of an on-operation time period and an off-operation time period of the compressor12is controlled to control the refrigerant discharge rate of the compressor12.

Furthermore, in a case where an electric compressor is used, the refrigerant discharge rate of the electric compressor12can be controlled by controlling the rpm of the electric compressor12.

(8) In the first to tenth embodiments, if the ejector14is a flow rate variable ejector, in which a cross sectional area of the refrigerant flow passage of the nozzle portion14ais variable based on the sensed superheat at the outlet of the first evaporator15, the discharged refrigerant pressure (the flow rate of the refrigerant to be drawn into the ejector14) can be controlled.

Thus, in each of the multiple evaporator operating modes of the second embodiment, the first and second evaporator operating mode and the first to third evaporator operating mode of the third embodiment, the flow rate of the refrigerant, which flows through the second evaporator18, can be more precisely controlled.

(9) In the first to tenth embodiments, the multiple evaporators (e.g., the first and second evaporators15,18) can be integrally assembled as a single unit.

(10) In the eleventh and twelfth embodiments, the present invention is implemented in the vehicle air conditioning system. However, the present invention is not limited to the vehicle air conditioning system. For example, the present invention can be implemented in any other vapor compression type cycle, such as a heat pump cycle of a water heater. Furthermore, in the eleventh and twelfth embodiments, the first and second evaporators104,106have the two different refrigeration capacities, respectively. Alternatively, three or more evaporators can be provided to have three or more different refrigeration capacities.

Furthermore, a receiver can be arranged downstream of the radiator102in the eleventh and twelfth embodiments. Also, a fixed ejector, in which the second metering portion S2of the nozzle portion131is stationary, may be used in place of the ejector103of the eleventh and twelfth embodiments. Furthermore, in the eleventh and twelfth embodiments, the two evaporators104,106of different refrigeration capacities are separately constructed. Alternatively, these evaporators104,106can be formed integrally, as discussed in the following embodiments.

THIRTEENTH EMBODIMENT

FIGS. 19 and 20show a thirteenth embodiment of the present invention. Specifically,FIG. 19shows an exemplary case where a vapor compression cycle210of the thirteenth embodiment is applied to a vehicular refrigeration cycle. In the cycle210of the present embodiment, a compressor211, which draws and compresses refrigerant, is driven by a vehicle drive engine (not shown) through, for example, a solenoid clutch212, a belt and the like.

The compressor211may be a variable displacement compressor or a fixed displacement compressor. In the case of the variable displacement compressor, a refrigerant discharge rate is adjusted by changing its displacement. In the case of the fixed displacement compressor, a refrigerant discharge rate is adjusted by varying its operating rate through repeated connection and disconnection of the solenoid clutch212. Furthermore, when an electric compressor is used as the compressor211, a refrigerant discharge rate can be adjusted by adjusting a rotational speed of an electric motor.

A radiator213is provided on a refrigerant outlet side of the compressor211. The radiator213exchanges heat between the high pressure refrigerant, which is discharged from the compressor211, and the external air (external air supplied from outside of the vehicle), which is blown toward the radiator213by a cooling fan (not shown), so that the high pressure refrigerant is cooled.

In a case where an ordinary fluorocarbon refrigerant is used as the refrigerant of the cycle210, the cycle210becomes a sub-critical pressure cycle, in which its high pressure does not exceed a critical pressure. Thus, the radiator213acts as a condenser, which condenses the refrigerant. In contrast, in a case where another type of refrigerant, such as a carbon dioxide (CO2) refrigerant, which has its high pressure exceeding the critical pressure, is used, the cycle210becomes a super-critical cycle. Thus, in such a case, the refrigerant radiates heat in a super-critical state without condensation of the refrigerant.

An ejector214is arranged on a downstream side of the radiator213in the refrigerant flow direction in the cycle210. The ejector214serves as a depressurizing means for depressurizing the refrigerant and is formed as a momentum-transporting pump, which performs fluid transportation by entraining action of discharged high velocity working fluid (see JIS Z 8126 Number 2.1.2.3).

The ejector214includes a nozzle portion214aand a refrigerant suction inlet214b. The nozzle portion214areduces a cross sectional area of the refrigerant passage, which conducts the high pressure refrigerant discharged from the radiator213, to isentropically depressurize and expand the high pressure refrigerant. The suction inlet214bis arranged in a space, in which a refrigerant outlet of the nozzle portion214ais located. The suction inlet214bdraws gas phase refrigerant supplied from a second evaporator218described below.

Furthermore, a mixing portion214cis provided on a downstream side of the nozzle portion214aand of the suction inlet214bin the refrigerant flow direction. The mixing portion214cmixes the high velocity refrigerant flow, which is outputted from the nozzle portion214a, with the drawn refrigerant, which is drawn through the suction inlet214b. A diffuser portion214d, which serves as a pressurizing portion, is arranged downstream of the mixing portion214cin the refrigerant flow direction. The diffuser portion214dis formed to progressively increase a cross sectional area of its refrigerant passage toward its downstream end. Thus, the diffuser portion214ddecelerates the refrigerant flow and increases the refrigerant pressure, i.e., the diffuser portion214dconverts the velocity energy of the refrigerant to the pressure energy.

A first evaporator215is connected to a downstream side of the diffuser portion214dof the ejector214, and a downstream side of the first evaporator215is connected to an inlet side of the compressor211.

A branched refrigerant passage (or simply referred to as a branched passage)216is branched from a branch point, which is located on an upstream side of the ejector214(an intermediate point between the radiator213and the ejector214), in the cycle210. A downstream side of the branched passage216is connected to the suction inlet214bof the ejector214. InFIG. 19, numeral Z indicates the branch point of the branched passage216.

A metering mechanism (or a flow rate control valve serving as a metering means)217is arranged in the branched passage216. The second evaporator218is arranged on a downstream side of the metering mechanism217. The metering mechanism217is a depressurizing means for adjusting a flow rate of the refrigerant supplied toward the second evaporator218. Specifically, the metering mechanism217may be a fixed choke or throttle, such as an orifice. Alternatively, the metering mechanism217may be an electric control valve, which is driven by an electric actuator to adjust a valve opening degree (a passage opening degree) of the control valve.

In the present embodiment, the two evaporators215,218are constructed integrally (assembled integrally or formed integrally) such that the two evaporators215,218are received in a single case219. A common electric blower220blows the air (air to be cooled) to an air passage in the case219, as indicated by an arrow A inFIG. 19, so that the blown air is cooled by the two evaporators215,218.

The cooled air, which is cooled by the two evaporators215,218, is supplied to a common subject cooling space221, so that the common subject cooling space221is cooled by the two evaporators215,218. Among the two evaporators215,218, the first evaporator215, which is connected to a main flow passage located on a downstream side of the ejector214, is arranged on an upstream side in the air flow A, and the second evaporator218, which is connected to the suction inlet214bof the ejector214, is arranged on a downstream side in the air flow A.

In a case where the cycle210of the present embodiment is applied to a refrigeration cycle of a vehicle air conditioning system, a passenger compartment of the vehicle becomes the subject cooling space221. In a case where the cycle210of the present embodiment is applied to a refrigeration cycle of a freezer and/or refrigerator (or simply indicated as “freezer/refrigerator”) vehicle, a freezer/refrigerator space of the freezer/refrigerator vehicle becomes the subject cooling space221.

Next, a specific example of the integrated structure of the two evaporators215,218will be described with reference toFIG. 20. In the example ofFIG. 20, the two evaporators215,218are integrated together as a single evaporator structure. Thus, the first evaporator215constitutes the upstream side section of the single evaporator structure, which is located on the upstream side in the air flow A. Furthermore, the second evaporator218constitutes the downstream side section of the single evaporator structure, which is located on the downstream side in the air flow A.

The structure of the first evaporator215and the structure of the second evaporator218are basically the same. Thus, each of the first and second evaporators215,218has a heat exchange core215a,218aand upper and lower tanks215b,215c,218b,218c. The upper and lower tanks215b,215c,218b,218care arranged on upper and lower sides, respectively, of the heat exchange core215a,218a.

The heat exchange core215a,218ahas a stack structure that includes a plurality of vertically extending tubes222and a plurality of fins223. Each fin223is connected between corresponding two of the tubes222. InFIG. 20, only the tubes222and the fins223of the heat exchange core215aof the first evaporator215, which is located on the upstream side in the air flow A, are depicted, while the tubes222and the fins223of the heat exchange core218aof the second evaporator218are not depicted for the sake of simplicity. However, it should be noted that the heat exchange cores215a,218ahave basically the same structure, as noted above.

The tubes222constitute the refrigerant passages and are made as generally planar tubes, each of which is planar, i.e., is generally flattened in the air flow direction A. The fins223are made as corrugated fins, each of which is formed by bending a thin plate material into a wavy form and is joined to planar outer surfaces of the corresponding tubes222to increase a heat transfer surface area for exchanging heat with the air.

The tubes222and the fins223are alternately stacked one after another in a left-right direction of the heat exchange core215a,218a. Two side plates215d,215e,218d,218eare arranged at opposed ends, respectively, of the heat exchange core215a,218a, which are opposed to each other in a stacking direction of the tubes222and the fins223(i.e., in the left-right direction of the heat exchange core215a,218a) to reinforce the heat exchange core215a,218a. The side plates215d,215e,218d,218eare connected to the left and right outermost corrugated fins223, respectively, and are also connected to the upper and lower tanks215b,215c,218b,218c.

The upper and lower tanks215b,215cof the first evaporator215forms a refrigerant passage space, which is independent from a refrigerant passage space, which is formed by the upper and lower tanks218b,218cof the second evaporator218. The upper and the lower tanks215b,215cof the first evaporator215have tube engaging holes (not shown), to which upper and lower ends of the tubes222of the heat exchange core215aare connected in such a manner that the upper and lower ends of the tubes222are communicated with interior spaces of the tanks215b,215c.

Similarly, the upper and the lower tanks218b,218cof the second evaporator218have tube engaging holes (not shown), to which upper and lower ends of the tubes222of the heat exchange core218aare connected in such a manner that the upper and lower ends of the tubes222are communicated with interior spaces of the tanks218b,218c.

In this way, each of the upper and lower tanks215b,215c,218b,218chas a role-of distributing the refrigerant flow to the tubes222of the corresponding heat exchange core215a,218aor a role of collecting the refrigerant flow from the tubes222.

The distribution and collection of the refrigerant flow by the tanks215b,215c,218b,218cwill be more specifically described with reference toFIG. 20. InFIG. 20, an inlet224, into which the low pressure refrigerant on the downstream side of the ejector214is supplied, is arranged at the left end of the lower tank215cof the first evaporator215, and an outlet225is arranged at the right end of the lower tank215c. A partition plate226is arranged generally in a longitudinal center of the interior space of the lower tank215c, which is centered in the longitudinal direction of the interior space of the lower tank215c(in the stacking direction of the tubes222and the fins223of the heat exchange core215a). The partition plate226divides the interior space of the lower tank215cinto a left region and a right region inFIG. 20.

Thus, the low pressure refrigerant, which is supplied from the inlet224into the left region of the interior of the lower tank215c, flows upward through a group of the left side tubes222of the heat exchange core215ain a direction of arrow “a” and then flows from the left side to the right side in the interior of the upper tank215bin a direction of arrow “b” inFIG. 20.

Then, the refrigerant, which is now located on the right region of the interior of the upper tank215b, flows downward through a group of right side tubes222of the heat exchange core215ain a direction of arrow “c” and enters the right region of the interior of the lower tank215cinFIG. 20. Then, the refrigerant is discharged in a direction of arrow “d” inFIG. 20from the outlet225, which is located in the right end of the lower tank215c, so that the refrigerant is directed toward an suctioning inlet side of the compressor211.

In contrast, in the second evaporator218, an inlet227, into which the low pressure refrigerant passed through the metering mechanism217of the branched passage216is supplied, is arranged at the right end of the upper tank218b. Furthermore, an outlet228is arranged at the left end of the upper tank218b. A partition plate229is arranged generally in a longitudinal center of the interior space of the upper tank218b, which is centered in the longitudinal direction of the interior space of the upper tank218b(in the stacking direction of the tubes222and the fins223of the heat exchange core218a). The partition plate229divides the interior space of the upper tank218binto a left region and a right region inFIG. 20.

Thus, the low pressure refrigerant, which is supplied from the inlet227into the right region of the interior of the upper tank218b, flows downward through a group of right side tubes222of the heat exchange core218ain a direction of arrow “e” and then flows from the right side to the left side in the interior of the lower tank218cin a direction of arrow “f” inFIG. 20.

Then, the refrigerant, which is now located on the left region of the interior of the lower tank218c, flows upward through a group of left side tubes222of the heat exchange core218ain a direction of arrow “g” and enters the left region of the interior of the upper tank218binFIG. 20. Then, the refrigerant is discharged in a direction of arrow “h” inFIG. 20from the outlet228, which is located in the left end of the upper tank218b, so that the refrigerant is directed toward the suction inlet214bside of the ejector214.

Next, the specific integral structure of the tubes222, the fins223and the tanks215b,215c,218b,218cof the two evaporators215,218will be described.

Separate arrangements of fins, which serve as the fins223, may be respectively provided to the two heat exchange cores215a,218a, which are arranged one after the other in the air flow A. Alternatively, a common single arrangement of fins, which serve as the fins223, may be provided commonly to the two heat exchange cores215a,218a.

Similarly, separate arrangements of tubes, which serve as the tubes222, may be respectively provided to the two heat exchange cores215a,218a, which are arranged one after the other in the air flow A. Alternatively, a common single arrangement of tubes, which serve as the tubes222, may be provided commonly to the two heat exchange cores215a,218a.

However, the tubes222of the first evaporator215and the tubes222of the second evaporator218need to form completely independent refrigerant passages, respectively. Thus, in the case where the integral single arrangement of tubes is used, the refrigerant passage of the first evaporator215and the refrigerant passage of the second evaporator218need to be separated from each other by corresponding partition walls provided in the tubes. In such a case, the refrigerant passages, which are defined by the tubes of the first evaporator215, need to be independently connected to the interiors of the upper and lower tanks215b,215cof the first evaporator215. Also, the refrigerant passages, which are defined by the tubes of the second evaporator218, need to be independently connected to the interiors of the upper and lower tanks218b,218cof the second evaporator218.

Also, the tanks215b,215c,218b,218cmay be independently formed. Alternatively, the two upper tanks215b,218bmay be constructed integrally, and the two lower tanks215c,218cmay be constructed integrally. However, even in this case too, the interior spaces of the upper tanks215b,218bneed to be formed independently of each other, and the interior spaces of the lower tanks215c,218cneed to be formed independently of each other.

In addition, the left and right side plates215d,215e,218d,218emay be formed independently from each other. Alternatively, the two left side plates215d,218dmay be formed integrally as a single plate, and the two right side plates215e,218emay be formed integrally as a single plate.

As discussed above, when the tubes222, the fins223, the tanks215b,215c,218b,218cand the side plates215d,215e,218d,218eof the first and second evaporators215,218are constructed as the integral structure, the number of components of the evaporators215,218can be reduced, and the manufacturing costs can be reduced.

A specific material of the tubes222, the fins223, the tanks215b,215c,218b,218cand the side plates215d,215e,218d,218eis preferably aluminum, which is the metal that exhibits the superior thermal conductivity and the superior solderability. However, the material is not limited to the aluminum and can be any other suitable material. When each component of the first and second evaporators215,218is made from the aluminum material, the first and second evaporators215,218can be joined together by the soldering.

In the present embodiment, after the assembling of the first and second evaporators215,218by the soldering, the ejector214is installed to the first and second evaporators215,218to integrate the ejector214with the first and second evaporators215,218.

As shown inFIG. 20, the ejector214is formed into an elongated cylindrical body, in which the nozzle portion214a, the mixing portion214cand the diffuser portion214dare arranged one after another along a straight line. Thus, in the present embodiment, the ejector214is assembled integrally to the lateral surfaces of the heat exchange cores215a,218ain such a manner that the longitudinal direction of the ejector214is made parallel to the lateral surfaces of the heat exchange cores215a,218a.

More specifically, the longitudinal direction of the ejector214is arranged parallel to the left side plates215d,218dof the heat exchange cores215a,218a, and the ejector214is installed to the left side plates215d,218d. Here, the ejector214is secured to the side plates215d,218dby a securing means (not shown), such as screws, metal spring clips or soldering.

With the above assembly structure of the ejector214, the outlet of the diffuser214dof the ejector214can be positioned close to the inlet224of the lower tank215c, and the suction inlet214bof the ejector214can be positioned close to the outlet228of the upper tank218b. Thus, both the refrigerant passage connection between the ejector214and the first evaporator215and the refrigerant passage connection between the ejector214and the second evaporator218can be made simple.

Furthermore, the longitudinal direction of the ejector214, which is made as the elongated cylindrical body, is arranged along the lateral surfaces of the heat exchange cores215a,218aof the first and second evaporators215,218, so that the ejector214will not protrude significantly outward from the first and second evaporators215,218. As a result, the entire size of the first and second evaporators215,218and the ejector214can be made compact.

Next, operation of the thirteenth embodiment will be described. When the compressor211is driven by the vehicle engine, the refrigerant is compressed in the compressor211. Then, the high temperature and high pressure refrigerant is discharged from the compressor211and is supplied to the radiator213. In the radiator213, the high temperature refrigerant is cooled by the external air and is thus condensed. In the branch point Z, the high pressure liquid state refrigerant, which is discharged from the radiator213, is divided into a refrigerant flow directed to the ejector214and a refrigerant flow directed to the branched passage216.

The refrigerant flow, which is supplied to the ejector214, is depressurized and is expanded at the nozzle portion214a. Thus, the pressure energy of the refrigerant is converted into the velocity energy at the nozzle portion214a, and thereby the refrigerant is discharged at the high velocity from outlet of the nozzle portion214a. Due to the decrease in the refrigerant pressure, the refrigerant (gas phase refrigerant), which has passed through the second evaporator218in the branched passage216, is drawn through the suction inlet214b.

The refrigerant, which is discharged from the nozzle portion214a, and the refrigerant, which is drawn through the suction inlet214b, are mixed in the mixing portion214clocated on the downstream side of the nozzle portion214aand are then supplied to the diffuser portion214d. In the diffuser portion214d, due to the increase in the passage cross sectional area, the velocity (expansion) energy is converted into the pressure energy, so that the pressure of the refrigerant increases.

The refrigerant, which is discharged from the diffuser portion214dof the ejector214, is supplied to the first evaporator215. In the first evaporator, while the refrigerant flows in the refrigerant flow path indicated by the arrows a-d inFIG. 20, the low temperature low pressure refrigerant absorbs the heat from the blown air, which is blown in the direction of arrow A, and is thus evaporated. After the evaporation, this gas phase refrigerant is drawn into and is compressed in the compressor211once again.

In contrast, the refrigerant flow, which is supplied to the branched passage216, is depressurized in the metering mechanism217and thus becomes the low pressure refrigerant. Then, the low pressure refrigerant is supplied to the second evaporator218. In the second evaporator218, while the refrigerant flows in the refrigerant flow path indicated by the arrows e-h inFIG. 20, the refrigerant absorbs the heat from the blown air, which is blown in the direction of arrow A. After the evaporation, this gas phase refrigerant is drawn into the ejector214through the suction inlet214b.

As discussed above, according to the present embodiment, the refrigerant on the downstream side of the diffuser portion214dof the ejector214can be supplied to the first evaporator215, and also the refrigerant in the branched passage216can be supplied to the second evaporator218through the metering mechanism217. Thus, the first and second evaporates215,218can simultaneously perform its cooling operation. Therefore, the cooled air, which is cooled by both the first and second evaporators215,218, can be discharged into the subject cooling space221to cool the subject cooling space221.

At this time, the refrigerant evaporation pressure of the first evaporator215is the pressure of the refrigerant after the increasing of the pressure in the diffuser214d, and the outlet of the second evaporator218is connected to the suction inlet214bof the ejector214. Thus, the lowest pressure right after the depressurization in the nozzle portion214acan be applied to the second evaporator218.

In this way, the refrigerant evaporation pressure (the refrigerant evaporation temperature) of the second evaporator218can be made lower than that of the first evaporator215. Furthermore, the first evaporator215, which has the higher refrigerant evaporation temperature, is arranged on the upstream side in the air flow direction A, and the second evaporator218, which has the lower refrigerant evaporation temperature, is arranged on the downstream side in the air flow direction A. Thus, the required temperature difference between the refrigerant evaporation temperature and the blown air temperature at the first evaporator215and the required temperature difference between the refrigerant evaporation temperature and the blown air temperature at the second evaporator218can be both satisfied.

As a result, the cooling performance of the first evaporator215and the cooling performance of the second evaporator218can be effectively achieved. Therefore, the cooling performance for cooling the common subject cooling space221can be effectively improved by the combination of the first and second evaporators215,218. Furthermore, the intake pressure of the compressor221is increased by the pressure increasing operation of the diffuser portion214d, so that the required drive force for driving the compressor211can be reduced.

Also, in the cycle210of the present embodiment, the branched passage216, which is branched at the branch point Z, is connected to the suction inlet214bof the ejector214, and the metering mechanism217and the second evaporator218are arranged in the branched passage216. Thus, the low pressure two-phase refrigerant of a gas and liquid mixture can be supplied to the second evaporator218through the branched passage216. Therefore, it is not required to provide the gas-liquid separator of, for example, Japanese Patent No. 33222263 (corresponding to U.S. Pat. Nos. 6,477,857 and 6,574,987).

In the case where the super-critical cycle, in which the above gas-liquid separator is provided, and the refrigerant, such as CO2, which has the high cycle pressure that exceeds the critical pressure, is used, when the operation of the cycle is stopped under the high external temperature, the low pressure side of the cycle also becomes the critical state in addition to the high pressure side.

Thus, the gas phase refrigerant and the liquid phase refrigerant cannot be separated by the gas-liquid separator at the time of restarting the cycle operation. Therefore, the high temperature refrigerant of the super-critical state, which is present in the gas-liquid separator, is supplied to the second evaporator218, so that the cooling performance of the second evaporator218is significantly reduced. In contrast, according to the present embodiment, the high pressure refrigerant is branched off on the upstream side of the ejector214, and this branched refrigerant is depressurized through the metering mechanism217to supply the low pressure refrigerant to the inlet side of the second evaporator218. As a result, the cooling performance of the second evaporator218can be quickly enabled even at the time of restarting the cycle operation.

Furthermore, in a sub-critical cycle (cycle having its high pressure without exceeding the critical pressure), which uses the ordinary fluorocarbon refrigerant, the pressure difference between the high pressure and the low pressure of the cycle becomes small in the small cycle thermal load condition, so that the input to the ejector214is reduced. In such a case, in the cycle recited in Japanese Patent No. 33222263, the refrigerant flow, which passes through the second evaporator218, depends only on the refrigerant drawing performance of the ejector214. Thus, when the decrease in the input of the ejector214occurs, the refrigerant drawing performance of the ejector214is reduced, and the refrigerant flow rate of the second evaporator218is reduced. Therefore, it is difficult to achieve the required cooling performance of the second evaporator218.

In contrast, according to the present embodiment, the high pressure refrigerant is branched on the upstream side of the ejector214, and this branched refrigerant is drawn into the suction inlet214bof the ejector214through the branched passage216. Thus, the branched passage216is connected in parallel with the ejector214.

Therefore, besides the refrigerant drawing performance of the ejector214, the refrigerant drawing performance and the refrigerant discharging performance of the compressor211can be utilized to supply the refrigerant in the branched passage216. In this way, even when the input of the ejector214is reduced to cause the reduction of the refrigerant drawing performance of the ejector214, the decrease in the refrigerant flow rate on the second evaporator218side can be alleviated in comparison to the cycle recited in Japanese Patent No. 33222263. Therefore, even in the low thermal load condition, the required cooling performance of the second evaporator can be more easily achieved.

Furthermore, the refrigerant flow rate on the second evaporator218side can be independently adjusted by the metering mechanism217without relying on the function of the ejector214. The flow rate of the refrigerant, which is supplied to the first evaporator215, can be adjusted through the control of the refrigerant discharging performance of the compressor211and the metering characteristics of the ejector214. Thus, the flow rate of the refrigerant to the first evaporator215and the flow rate of the refrigerant to the second evaporator218can be easily adjusted depending on the thermal load of the first evaporator215and the thermal load of the second evaporator218, respectively.

FOURTEENTH EMBODIMENT

In the thirteenth embodiment, the ejector214is assembled integrally to the lateral surfaces of the heat exchange cores215a,218ain such a manner that the longitudinal direction of the ejector214is made parallel to the lateral surfaces of the heat exchange cores215a,218a. In the fourteenth embodiment, as shown inFIG. 21, the ejector214is assembled integrally to the tanks215b,215c,218b,218cin such a manner that the longitudinal direction of the ejector214is made parallel to the tanks215b,215c,218b,218c.

More specifically, in the exemplary case ofFIG. 21, the ejector214is assembled integrally to the top surfaces of the upper tanks215b,218bin such a manner that the longitudinal direction of the ejector214is made parallel to the top surfaces of the upper tanks215b,218b. The securing means for securing the ejector214to the top surfaces of the upper tanks215b,218bmay be the same as that of the thirteenth embodiment.

Next, the refrigerant passage structures of the first and second evaporators215,218will be described. In the first evaporator215, the partition plate226is provided in the upper tank215bto divide the interior space of the upper tank215binto the left region and the right region inFIG. 21. The inlet224is arranged in the right region of the top surface of the upper tank215b, and the downstream side passage of the diffuser portion214dof the ejector214is connected to the inlet224. Furthermore, the outlet225is arranged in the lateral surface of the left region of the upper tank215b.

The refrigerant, which is supplied from the inlet224to the right region of the upper tank215b, passes through the right region of the heat exchange core215a, the lower tank215c, the left region of the heat exchange core215aand the left region of the upper tank215bin this order and is discharged from the outlet225toward the inlet of the compressor221, as indicated by arrows i, k, m and n inFIG. 21.

In the second evaporator215, similar to the thirteenth embodiment, the partition plate229is provided in the upper tank218bto divide the interior space of the upper tank218binto the left region and the right region inFIG. 21. The inlet227is arranged in the rear surface of the right region of the upper tank218b, and a connection pipe216a, which is arranged on the downstream side of the metering mechanism217of the branched passage216, is connected to the inlet227.

The refrigerant, which is supplied from the inlet227to the right region of the upper tank218b, passes through the right region of the heat exchange core218a, the lower tank218c, the left region of the heat exchange core218aand the left region of the upper tank218bin this order and is supplied to the suction inlet214bof the ejector214, as indicated by arrows p, q, r and s inFIG. 21.

In the fourteenth embodiment, the location of the ejector214and the refrigerant passage structures of the first and second evaporators215,218are different from those of the thirteenth embodiment. However, the arrangement of the first and second evaporators215,218with respect to the air flow direction A and the passage structure of the cycle210are the same as those of the thirteenth embodiment. Thus, advantages similar to those of the thirteenth embodiment are also achieved in the fourteenth embodiment.

FIFTEENTH EMBODIMENT

In the thirteenth and fourteenth embodiments, the branched passage216, which is branched on the upstream side of the ejector214and is connected to the suction inlet214bof the ejector214, is provided, and the second evaporator218is provided in the branched passage216. However, in the fifteenth embodiment, the branched passage216is not provided.

That is, in the fifteenth embodiment, as shown inFIG. 22, a gas-liquid separator230is provided on the downstream side of the first evaporator215to separates the refrigerant of the gas and liquid mixture into the gas phase refrigerant and the liquid phase refrigerant. A gas phase refrigerant outlet of the gas-liquid separator230is connected to the inlet of the compressor211, and a liquid phase refrigerant outlet of the gas-liquid separator230is connected to the suction inlet214bof the ejector214through a branched refrigerant passage (or simply referred to as a branched passage)231. The metering mechanism217and the second evaporator218are provided in the branched passage231.

The arrangement of the first and second evaporators215,218with respect to the air flow direction A is the same as that of the thirteenth and fourteenth embodiments. Thus, the first evaporator215, which has the higher refrigerant evaporation temperature, is arranged on the upstream side in the air flow direction, and the second evaporator218, which has the lower refrigerant evaporation temperature, is arranged on the downstream side in the air flow direction A. The first and second evaporators215,218are integrated by the structure shown inFIGS. 20or21.

Even in the fifteenth embodiment, the cooling performance for cooling the subject cooling space221is advantageously improved by the combination of the first and second evaporators215,218, which have the different refrigerant evaporation temperatures.

SIXTEENTH EMBODIMENT

In the sixteenth embodiment, the cycle structure of the thirteenth or fourteenth embodiment is modified. Specifically, as shown inFIG. 23, the cycle structure of the sixteenth embodiment includes first and second low pressure passages232,233, which are branched from the downstream side of the ejector214and are connected to the input side of the compressor211. Furthermore, the first and second low pressure passages232,233are arranged in parallel. The cycle structure further includes first and second branched refrigerant passages (or simply referred to as first and second branched passages)216c,216d, which are branched at the upstream side of the ejector214and are connected to the suction inlet214bof the ejector214.

The two first evaporators215f,215gare provided in the first and second low pressure passages232,233, respectively, on the downstream side of the ejector14. Two metering mechanisms217a,217bare arranged in the first and second branched passages216c,216d, respectively, and the two second evaporators218f,218gare provided on the downstream side of the metering mechanisms217a,217b, respectively.

In the sixteenth embodiment, the first evaporator215fand the second evaporator218fare constructed integrally (assembled integrally or formed integrally) and are received in a single common case219a. A common electric blower (not shown but corresponding to the blower220ofFIG. 19) blows the air (air to be cooled) to an air passage in the case219a, as indicated by an arrow A1inFIG. 23, so that the blown air is cooled by the two evaporators215f,218f.

Similarly, the first evaporator215gand the second evaporator218gare constructed integrally, assembled integrally or are formed integrally and are received in a single common case219b. A common electric blower (not shown but corresponding to the blower220ofFIG. 19) blows the air (air to be cooled) to an air passage in the case219b, as indicated by an arrow A2inFIG. 23, so that the blown air is cooled by the two evaporators215g,218g.

The integration of the first evaporator215fand the second evaporator218fand the integration of the first evaporator215gand the second evaporator218gcan be implemented by the structure shown inFIG. 20orFIG. 21. The ejector214may be integrally assembled to any desired one of the integrated structure of the first and second evaporators215f,218fand the integrated structure of the first and second evaporators215g,218g.

The cool air, which is cooled by the two evaporators215f,218fin the case219a, is supplied into the common subject cooling space (not shown), so that the common subject cooling space is cooled by the two evaporators215f,218f.

Similarly, the cool air, which is cooled by the two evaporators215g,218gin the case219b, is supplied into the common subject cooling space (not shown), so that the common subject cooling space is cooled by the two evaporators215g,218g.

The subject cooling space of the case219aand the subject cooling space of the case219bare formed independently from each other. The subject cooling space of the case219amay be, for example, the passenger compartment of the vehicle, and the subject cooling space of the case219bmay be, for example, the freezer/refrigerator space of the freezer/refrigerator vehicle.

In the sixteenth embodiment, each first evaporator215f,215g, which has the higher refrigerant evaporation temperature, is arranged on the upstream side in the corresponding air flow direction A1, A2, and each second evaporator218f,218g, which has the lower refrigerant evaporation temperature, is arranged on the downstream side in the corresponding air flow direction A1, A2.

The thirteenth to sixteenth embodiments can be modified in various ways, as discussed below.

(1) In the cycle210of the thirteenth and sixteenth embodiments shown inFIGS. 19 and 23, there is not provided the gas-liquid separator, which separates the refrigerant of gas and liquid mixture into the gas phase refrigerant and the liquid phase refrigerant and accumulates the excessive refrigerant as the liquid refrigerant. However, for example, a gas-liquid separator (receiver), which separates the refrigerant of gas and liquid mixture into gas phase refrigerant and liquid phase refrigerant and accumulates the liquid phase refrigerant, may be provided at the outlet side of the radiator213, so that the liquid phase refrigerant is supplied from the gas-liquid separator to the ejector214. Furthermore, a gas-liquid separator (accumulator), which separates the refrigerant of gas-liquid mixture into gas phase refrigerant and liquid phase refrigerant and accumulates excessive refrigerant as the liquid phase refrigerant, may be provided at the inlet side of the compressor211, so that the gas phase refrigerant is supplied from the gas-liquid separator to the inlet of the compressor211.

(2) In each of the above embodiments, the vehicular refrigeration cycle is described. However, the present invention is not limited to the vehicular refrigeration cycle and can be equally applicable to a stationary refrigeration cycle, which is settled stationary.

(3) In each of the above embodiments, the type of refrigerant is not specified. However, it should be noted that the refrigerant of the above embodiments may be a fluorocarbon refrigerant (including chlorofluorocarbon refrigerant), an alternative for the chlorofluorocarbon refrigerant, such as hydrocarbon (HC) refrigerant, or carbon dioxide (CO2), which can be used in any one of the vapor compression type super critical cycle and the vapor compression type sub-critical cycle.

Here, it should be noted that the chlorofluorocarbon is a generic name of an organic compound composed of carbon, fluorine, chlorine and hydrogen and is widely used as the refrigerant. Furthermore, the fluorocarbon refrigerants include hydrochlorofluorocarbon (HCFC) refrigerant, hydrofluorocarbon (HFC) refrigerant, which do not destroy the ozone layer and thus are called as substitutes for the chlorofluorocarbon.

The hydrocarbon (HC) refrigerant is a refrigerant that includes hydrogen and carbon and is found in nature. The HC refrigerants include R600a (isobutene), R290 (propane) and the like.

(4) In each of the above embodiments, a flow rate variable type ejector, in which a cross sectional area of the refrigerant passage of the nozzle portion214a, i.e., a refrigerant flow rate in the nozzle portion214ais adjusted, may be used as the ejector214.

(5) In contrast to each of the above embodiments, the first evaporator215,215f,215g, which has the higher refrigerant evaporation temperature, may be arranged on the downstream side in the air flow direction A, A1, A2, and the second evaporator218,218f,218g, which has the lower refrigerant evaporation temperature, may be arranged on the upstream side in the air flow direction A1, A2.

(6) With reference toFIG. 24, the first evaporator215and the second evaporator218may be connected to each other by a refrigerant pipeline340through the ejector214. More specifically, the outlet of the second evaporator218may be connected to the suction inlet214bof the ejector214by a portion of the pipeline340, and the outlet of the diffuser portion214dof the ejector214may be connected to the inlet of the first evaporator215by another portion of the pipeline340. In this instance, as shown inFIG. 24, the first and second evaporators215,218may be constructed integrally in such a manner that a predetermined space is provided between the first evaporator215and the second evaporator218, and the refrigerant pipeline340integrally connects between the first and second evaporators215,218while limiting disassembly between the first and second evaporators215,218.

Additional advantages and modifications will readily occur to those skilled in the art. The invention in its broader terms is therefore not limited to the specific details, representative apparatus, and illustrative examples shown and described. Furthermore, it should be noted that the feature(s) of one of the above-described embodiments or modification(s) can be combined with the feature(s) of any other one of the above-described embodiments or modification(s).