Patent Publication Number: US-2009232665-A1

Title: Ejector

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application is based on Japanese Patent Applications No. 2008-062142 filed on Mar. 12, 2008, No. 2008-062143 filed on Mar. 12, 2008, No. 2008-135077 filed on May 23, 2008, No. 2008-135076 filed on May 23, 2008, and No. 2009-010645 filed on Jan. 21, 2009, the contents of which are incorporated herein by reference in its entirety. 
     FIELD OF THE INVENTION 
     The present invention relates to an ejector configured to draw a fluid by a jet flow of a high-speed fluid jetted from a nozzle. For example, the ejector can be suitably used for a refrigeration cycle device. 
     BACKGROUND OF THE INVENTION 
     Conventionally, an ejector is known, which includes a nozzle for decompressing and expanding a high-pressure fluid, and is configured to draw a fluid from a fluid suction port by a suction action of a jet flow of a high-speed fluid jetted from the nozzle. In the ejector, the jet fluid from the nozzle and the suction fluid from the fluid suction port are mixed in a mixing portion, and the pressure of the mixed fluid is increased in a diffuser portion by converting the kinetic energy of the mixed fluid to the pressure energy of the mixed fluid. Therefore, the pressure of the fluid flowing out of the outlet of the ejector is increased more than the pressure of the suction fluid. 
     In an ejector described in JP 2004-340136A (corresponding to US 2004/0206111 A1), a passage sectional area at an inlet side of a suction passage through which a suction fluid introduced from a fluid suction port flows into a mixing portion of the ejector is set equal to or larger than a passage sectional area of the fluid suction port. Therefore, the pressure loss, caused when the suction fluid is drawn from the fluid suction port, can be reduced, and the flow amount of the suction fluid flowing from the fluid suction port can be increased, thereby improving ejector efficiency ηe that is an energy converting efficiency in the ejector. 
     In an ejector for a refrigeration cycle device described in JP 2003-14318A (corresponding to US 2002/0000095A1), an expanding angle of a passage wall surface of a diffuser portion is suitably set in an axial section including the center axial of a nozzle so that a pressurizing amount in the diffuser portion is increased, thereby improving the ejector efficiency ηe. 
     In another ejector described in JP 2004-116807A, a passage wall surface of a diffuser portion is formed into a smoothly covered line in an axial section including the center axial of a nozzle so that an energy loss such as a scroll flow loss in the diffuser portion can be restricted, thereby improving the ejector efficiency ηe. 
     The ejector efficiency ηe is defined as in the following formula (F1). 
         e =(1+Ge/ Gnoz )×(Δ P/ ρ)/Δ i    (F1) 
     Here, Ge is the flow amount of the suction fluid, Gnoz is the flow amount of the jet fluid, ΔP is the pressurizing amount in the diffuser portion, ρ is the density of the suction fluid, and Δi is the enthalpy difference between the inlet and the outlet of the nozzle. 
     However, JP 2004-340136A does not describe regarding the pressure loss on a downstream side in the suction passage downstream of the fluid suction port. If the pressure loss in the suction passage changes, the flow amount of the suction fluid or the flow velocity of the fluid flowing into the mixing portion through the suction passage is changed. In addition, when the fluid flowing through the mixing portion and the diffuser portion is in a gas-liquid two-phase state, the inertial force becomes different in the gas fluid and the liquid fluid due to the density difference between the gas fluid and the liquid fluid, and thereby it is difficult to uniformly mix the jet fluid and the suction fluid in the mixing portion of the ejector. 
     Thus, in the diffuser portion of the ejector, the kinetic energy of the fluid is converted to the pressure energy in an inhomogeneous state, and thereby the ejector efficiency ηe cannot be sufficiently improved. Here, the inhomogeneous state means a state other than a homogeneous state that includes a complete gas state, a complete liquid state and a homogeneously mixed state in which the gas fluid and the liquid fluid are homogeneously mixed with approximately the same flow velocity. In an example of the inhomogeneously mixed sate of the gas fluid and the liquid fluid, the flow velocity of the gas fluid is different from the flow velocity of the liquid fluid. 
     Furthermore, in JP 2003-14318A or JP 2004-116807A, the ejector is configured to improve the ejector efficiency ηe, in a case where the fluid of the homogeneous state passes through the mixing portion and the diffuser portion of the ejector. Actually, it is difficult for the gas-liquid two-phase fluid passing through the mixing portion and the diffuser portion of the ejector to be in the homogeneous state. Accordingly, when gas-liquid two-phase refrigerant passes through the mixing portion and the diffuser portion in the ejector, it is difficult to sufficiently improve the ejector efficiency ηe. 
     SUMMARY OF THE INVENTION 
     In view of the foregoing problems, it is an object of the present invention to sufficiently improve the ejector efficiency ηe in an ejector having a mixing and pressurizing portion in which the kinetic energy of a gas-liquid two-phase fluid is converted to the pressure energy thereof. 
     It is another object of the present invention to provide an ejector provided with a suction passage which is configured to improve the ejector efficiency ηe. 
     The following aspects of the present invention are devised by the inventors of the present application based on the following experiments and studies. An ejector recovers the energy lost in decompression and expansion by decompressing and expanding a fluid in iso-entropy at a nozzle, and converts the recovered energy (recovery energy) to the pressure energy, so as to improve the ejector efficiency ηe. 
     If it is possible for all the recovery energy to be converted to the pressure energy, the ejector efficiency ηe will be made maximum. The inventors of the present application examined and studied in detail regarding the recovery energy used actually in the ejector. That is, the energy capable of being used for pressurizing the fluid, among the recovery energy, is studied. 
       FIG. 28  shows results examined and studied by the inventors of the present application. In an ejector of a comparative example having a mixing portion and a diffuser portion shown in  FIG. 29 , a total recovery energy at an inlet of a mixing portion can be divided into E 1  to E 4  as shown in  FIG. 28 . In  FIG. 28 , E 1  indicates the energy used for pressurizing, E 2  indicates a remaining kinetic energy without being used, E 3  indicates energy transmission loss, and E 4  indicates the other loss. As shown from  FIG. 28 , the energy E 1  used for pressurizing is about  20 % of the total recovery energy, and the other energy E 2 , E 3 , E 4  is not used for pressurizing. The remaining kinetic energy E 2  is remained as a flow velocity of the fluid flowing out of the diffuser portion of the ejector without being converted to the pressure energy. 
     The energy transmission loss E 3  includes the energy transmission loss caused by transmitting the kinetic energy of the liquid fluid to the gas fluid, while the liquid fluid and the gas refrigerant pass through the diffuser portion of the ejector, for example. As shown in  FIG. 28 , the ratio of the energy transmission loss E 3 , in the energy E 2 , E 3  and E 4  without being used for pressurizing, is relatively large, as compared with the energy E 1  used for pressurizing. 
     The inventors of the present application studied regarding the reduction of the energy transmission loss E 3  between the gas fluid and the liquid fluid. When the energy transmission loss E 3  between the gas fluid and the liquid fluid is reduced and is used for the pressurizing, the ejector efficiency ηe can be effectively improved. Thus, the inventors performed experiments for effectively transmitting energy from the liquid fluid having a high flow velocity than that of the gas fluid, to the gas fluid. 
     In a case of a free fall rigid body, the flow velocity in a vertical downward direction is increased by acceleration of gravity. Then, the flow velocity of the free fall rigid body is reached to a certain terminal velocity in accordance with a balance with resistance received from the circumference air. 
     That is, the flow velocity of the free fall rigid body is not increased more than the terminal velocity after reaching the terminal velocity. Therefore, the flow velocity of the free fall rigid body becomes maximum when reaching to the terminal velocity. It means that the kinetic energy of the rigid body can be rapidly transmitted to the circumference air when the rigid body rapidly reaches to the terminal velocity. In  FIG. 29 , the liquid fluid grain (i.e., virtual liquid particle) passing through the diffuser portion is supposed as the rigid body, and the gas fluid passing through the diffuser portion is supposed as the circumference air. In the supposed state of  FIG. 29 , the inventors of the present application studied regarding an effective energy transmission between the liquid fluid and the gas fluid passing through the diffuser portion. 
     The upper part of  FIG. 29  is a graph showing variation in velocity of the gas fluid and velocity of the liquid fluid within an ejector. The solid line LA 1  shows a variation in the liquid fluid (e.g., liquid refrigerant) in the ejector of the comparison example, and the solid line GA 1  shows a variation in the gas fluid (e.g., gas refrigerant) in the ejector of the comparison example. As shown in the solid lines LA 1  and GA 1  of  FIG. 29 , the flow velocity of the gas fluid is greatly faster than that of the liquid fluid in a nozzle of the ejector in the comparison example by the difference in the inertial force due to the density difference between the gas fluid and the liquid fluid. Thus, in the mixed fluid of the jet fluid and the suction fluid flowing into a mixing portion of the ejector, the flow velocity of the gas fluid becomes faster than the flow velocity of the liquid refrigerant. 
     The grains of the liquid fluid flowing into the mixing portion are accelerated together with the circumference gas fluid, and then the flow velocity of the grains of the liquid fluid becomes equal to the flow velocity of the gas fluid. After the flow velocity of the grains of the liquid fluid becomes equal to the flow velocity of the gas fluid, the flow velocity of the liquid fluid is not more accelerated, and reaches to the terminal velocity. 
     The flow velocity of the grain of the liquid fluid after reaching to the terminal velocity is reduced while applying a force corresponding to the resistance force to the circumferential gas fluid as a reaction force. At this time, the kinetic amount is transmitted from the grains of the liquid fluid to the gas fluid, and the total value of impulses applied from the grains of the liquid fluid to the gas fluid becomes the pressurizing amount (pressure energy) of the gas fluid. 
     Accordingly, if the grains of the liquid fluid flowing into the mixing portion of the ejector are rapidly reached to the terminal velocity, the kinetic energy included in the liquid fluid can be rapidly transmitted to the gas fluid. Thus, after the flow velocity of the liquid fluid reaches to the terminal velocity, the kinetic energy of the liquid fluid can be effectively transmitted to the gas fluid. Furthermore, when the terminal velocity itself of the grains of the liquid fluid is increased, the pressurizing amount of the gas fluid can be increased, thereby improving the ejector efficiency ηe. 
     In  FIG. 29 , the chain line LA 2  indicates a variation in the flow velocity of the liquid fluid of an ejector according to an example of the present invention, and the chain line GA 2  indicates a variation in the flow velocity of the gas fluid of the ejector according to the example of the present invention. As shown by the chain lines LA 2  and GA 2  in  FIG. 29 , when the flow velocity of the gas fluid flowing into the mixing portion is increased, the terminal velocity of the grains of the liquid fluid can be increased, as compared with the comparison example shown by the solid lines LA 1  and GA 1 . Thus, in the example of the present invention shown by the chain lines LA 2  and GA 2  in  FIG. 29 , because a large amount of the kinetic energy can be converted to the pressure energy, the energy transmission loss between the gas fluid and the liquid fluid can be effectively reduced, thereby significantly improving the ejector efficiency ηe. 
     According to an aspect of the present invention, an ejector includes a nozzle configured to decompress and expand a fluid in any one state of a gas-liquid two-phase state, a liquid state and a super-critical state, and a body portion in which the nozzle is disposed. The body portion has a fluid suction port from which a fluid is drawn by a jet flow of the fluid jetted from a jet port of the nozzle, and a mixing and pressurizing portion in which the fluid jetted from the jet port of the nozzle and the fluid drawn from the fluid suction port are mixed and kinetic energy of the mixed fluid in a gas-liquid two-phase state is converted to pressure energy thereof. The ejector is provided with a suction passage through which the fluid drawn from the fluid suction port flows into an inlet of the mixing and pressurizing portion, and a fluid passage area of the suction passage is configured to be changed such that the fluid drawn from the fluid suction port is decompressed in the suction passage substantially in iso-entropy. 
     Accordingly, the energy loss while the suction fluid passes through the suction passage can be reduced. Thus, the flow velocity of the fluid flowing into the mixing and pressurizing portion from the suction passage can be increased, thereby increasing the flow velocity of the gas fluid flowing into the mixing and pressurizing portion. As a result, the terminal velocity of grains of the liquid fluid flowing into the mixing and pressurizing portion can be increased, and the pressurizing amount in the gas fluid can be increased in the ejector even when the kinetic energy of the gas-liquid two-phase fluid is converted to the pressure energy thereof in the mixing and pressurizing portion. Therefore, the ejector efficiency can be effectively improved. 
     According to another aspect of the present invention, an ejector includes a nozzle configured to decompress and expand a fluid in any one state of a gas-liquid two-phase state, a liquid state and a super-critical state, and a body portion in which the nozzle is disposed. The body portion has a fluid suction port from which a fluid is drawn by a jet flow of the fluid jetted from a jet port of the nozzle, and a mixing and pressurizing portion in which the fluid jetted from the jet port of the nozzle and the fluid drawn from the fluid suction port are mixed and kinetic energy of the mixed fluid in a gas-liquid two-phase state is converted to pressure energy thereof. The ejector is provided with a suction passage through which the fluid drawn from the fluid suction port flows into an inlet of the mixing and pressurizing portion. In the ejector, a fluid passage area of the suction passage is configured to be changed such that a flow velocity of the fluid flowing into the mixing and pressurizing portion from the suction passage is substantially equal to a flow velocity of the fluid flowing from the jet port of the nozzle into the mixing and pressurizing portion. As a result, the terminal velocity of grains of the liquid fluid flowing into the mixing and pressurizing portion can be increased, and the pressurizing amount in the gas fluid can be increased in the ejector even when the kinetic energy of the gas-liquid two-phase fluid is converted to the pressure energy thereof in the mixing and pressurizing portion. Therefore, the ejector efficiency can be effectively improved. Here, the meaning of “substantially equal” includes that the flow velocity of the fluid flowing into the mixing and pressurizing portion from the suction passage completely corresponds to or slightly different from the flow velocity of the fluid flowing from the jet port of the nozzle into the mixing and pressurizing portion. 
     According to another aspect of the present invention, an ejector includes a nozzle configured to decompress and expand a fluid in any one state of a gas-liquid two-phase state, a liquid state and a super-critical state, and a body portion in which the nozzle is disposed. The body portion has a fluid suction port from which a fluid is drawn by a jet flow of the fluid jetted from a jet port of the nozzle, and a mixing and pressurizing portion in which the fluid jetted from the jet port of the nozzle and the fluid drawn from the fluid suction port are mixed and kinetic energy of the mixed fluid in a gas-liquid two-phase state is converted to pressure energy thereof. The ejector is provided with a suction passage through which the fluid drawn from the fluid suction port flows into an inlet of the mixing and pressurizing portion. Furthermore, a fluid passage area of the suction passage is configured to be changed such that a flow velocity of the fluid flowing into the mixing and pressurizing portion from the suction passage is equal to or larger than a sound velocity. As a result, the terminal velocity of grains of the liquid fluid flowing into the mixing and pressurizing portion can be increased, and the pressurizing amount in the gas fluid can be increased in the ejector even when the kinetic energy of the gas-liquid two-phase fluid is converted to the pressure energy thereof in the mixing and pressurizing portion. 
     In any one aspect of the present invention, the fluid passage area of the suction passage may be gradually reduced toward downstream in a flow direction of the fluid flowing in the suction passage. In this case, a reduce degree of the fluid passage area at an inlet side of the suction passage may be larger than a reduce degree of the fluid passage area at an outlet side of the suction passage. 
     Alternatively, the fluid passage area of the suction passage at an inlet side of the suction passage may be gradually reduced toward downstream in the flow direction of the fluid flowing in the suction passage, and the fluid passage area of the suction passage at an outlet side of the suction passage may be gradually increased toward downstream in the flow direction of the fluid flowing in the suction passage. 
     The suction passage may be provided between an outer peripheral surface of the nozzle and an inner peripheral surface of the body portion, or may be configured by another nozzle to be provided therein. Alternatively, the nozzle and the suction passage may be configured, such that an enthalpy difference (ΔH) between enthalpy of the fluid at an inlet of the nozzle and enthalpy of the fluid at the jet port of the nozzle is equal to or larger than an enthalpy difference (Δh) between enthalpy of the fluid at the inlet of the suction passage and enthalpy of the fluid at the outlet of the suction passage. 
     According to another aspect of the present invention, an ejector includes a nozzle configured to decompress and expand a fluid in any one state of a gas-liquid two-phase state, a liquid state and a super-critical state, and a body portion in which the nozzle is disposed. The body portion has a fluid suction port from which a fluid is drawn by a jet flow of the fluid jetted from a jet port of the nozzle, and a mixing and pressurizing portion in which the fluid jetted from the jet port of the nozzle and the fluid drawn from the fluid suction port are mixed and kinetic energy of the mixed fluid in a gas-liquid two-phase state is converted to pressure energy thereof. The mixing and pressurizing portion is configured by a straight portion extending from the inlet of the mixing and pressurizing portion in a range, and an expanding portion extending from a downstream end of the straight portion to the outlet of the mixing and pressurizing portion. The straight portion is cylindrical passage having a constant passage area in its entire range, and the expending portion is configured such that a passage sectional area of the expanding portion is gradually increased toward downstream in a flow direction of the fluid. As a result, the terminal velocity of grains of the liquid fluid flowing into the mixing and pressurizing portion can be increased, and the pressurizing amount in the gas fluid can be increased in the ejector even when the kinetic energy of the gas-liquid two-phase fluid is converted to the pressure energy thereof in the mixing and pressurizing portion. 
     For example, the range of the straight portion may be set such that the flow velocities of gas fluid and liquid fluid within the fluid flowing into the mixing and pressurizing portion become equal to each other in the range. Alternatively, when a length of the straight portion in an axial direction of the nozzle is L 1  and a length from the inlet of the mixing and pressurizing portion to the outlet of the mixing and pressurizing portion in the axial direction is L 2 , the mixing and pressurizing portion is configured such that 0&lt;L 1 /L 2 ≦0.4. Furthermore, the mixing and pressurizing portion may be configured such that the fluid is pressurized in iso-entropy in the mixing and pressurizing portion. 
     In the ejector, a sectional shape of a wall surface of the expanding portion in a section including an axial line of the nozzle may be a straight line or a curved line. Alternatively, the sectional shape of the wall surface of the expanding portion in a section including the axial line of the nozzle may be formed by combining plural straight lines or may be formed by combining at least a straight line and a curved line. Alternatively, an expanding degree of the expanding portion at an inlet side of the expanding portion may be larger than an expanding degree of the expanding portion at an outlet side of the expanding portion. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Additional objects and advantages of the present invention will be more readily apparent from the following detailed description of preferred embodiments when taken together with the accompanying drawings. In which: 
         FIG. 1  is a schematic diagram showing a refrigeration cycle device having an ejector according to a first embodiment of the present invention; 
         FIG. 2A  is an axial sectional view of the ejector including an axial line of a nozzle according to the first embodiment,  FIG. 2B  is a cross-sectional view taken along the line IIB-IIB of  FIG. 2A , and  FIG. 2C  is a cross-sectional view taken along the line IIC-IIC of  FIG. 2A ; 
         FIG. 3  is a graph showing a variation in a ratio of a refrigerant passage sectional area of a suction passage to a refrigerant passage sectional area at an inlet of the suction passage, in the ejector according to the first embodiment; 
         FIG. 4  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion of the ejector according to the first embodiment; 
         FIG. 5A  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of the refrigeration cycle device according to the first embodiment, and  FIG. 5B  is an enlarged view showing the part VB in  FIG. 5A ; 
         FIG. 6  is a graph showing variations in the flow velocity of gas refrigerant and the flow velocity of liquid refrigerant in the ejector of the first embodiment and in an ejector of a comparison example; 
         FIG. 7A  is a graph showing variations in a flow velocity of refrigerant and a pressurizing amount (ΔP) in the ejector according to the first embodiment, and  FIG. 7B  is a graph showing variations in a flow velocity of refrigerant and a pressurizing amount (ΔP) in the ejector according to a comparison example; 
         FIG. 8  is a graph showing an energy amount (E 1 ) to be used for pressurizing, a remain kinetic energy (E 2 ), an energy transmission loss (E 3 ) and the other loss (E 4 ), according to the first embodiment and the comparison example; 
         FIG. 9  is a graph showing a variation in a ratio of a refrigerant passage sectional area of a suction passage to a refrigerant passage sectional area at an inlet of the suction passage, in an ejector according to a second embodiment of the present invention; 
         FIG. 10  is a graph showing a variation in a ratio of a refrigerant passage sectional area of a suction passage to a refrigerant passage sectional area at an inlet of the suction passage, in an ejector according to a third embodiment of the present invention; 
         FIG. 11  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion in an ejector according to a fourth embodiment of the present invention; 
         FIG. 12  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion in an ejector according to a fifth embodiment of the present invention; 
         FIG. 13  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion in an ejector according to a sixth embodiment of the present invention; 
         FIG. 14  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion in an ejector according to a seventh embodiment of the present invention; 
         FIG. 15  is a schematic diagram showing a passage configuration of a mixing and pressurizing portion in an ejector according to an eighth embodiment of the present invention; 
         FIG. 16  is an axial sectional view showing an ejector according to a ninth embodiment of the present invention; 
         FIG. 17  is an axial sectional view showing an ejector according to a tenth embodiment of the present invention; 
         FIG. 18  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of a refrigeration cycle device according to an eleventh embodiment of the present invention; 
         FIG. 19  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of a refrigeration cycle device according to a twelfth embodiment of the present invention; 
         FIG. 20  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of a refrigeration cycle device according to a thirteenth embodiment of the present invention; 
         FIG. 21  is a Mollier diagram showing another refrigerant state in the refrigerant cycle of the refrigeration cycle device according to the thirteenth embodiment of the present invention; 
         FIG. 22  is a schematic diagram showing a refrigeration cycle device having an ejector according to a fourteenth embodiment of the present invention; 
         FIG. 23  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of the refrigeration cycle device according to the fourteenth embodiment of the present invention; 
         FIG. 24  is a schematic diagram showing a refrigeration cycle device having an ejector according to a fifteenth embodiment of the present invention; 
         FIG. 25  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of the refrigeration cycle device according to the fifteenth embodiment of the present invention; 
         FIG. 26  is a schematic diagram showing a refrigeration cycle device having an ejector according to another embodiment of the present invention; 
         FIG. 27A  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of a refrigeration cycle device according to another embodiment of the present invention, and  FIG. 27B  is a Mollier diagram showing a refrigerant state in a refrigerant cycle of a refrigeration cycle device according to another embodiment of the present invention; 
         FIG. 28  is a graph showing energy division in a recovery energy at an inlet of a mixing portion of an ejector in a comparison example; and 
         FIG. 29  is a graph showing experimental results in velocity distribution of gas fluid and liquid fluid in an ejector. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     First Embodiment 
     A first embodiment of the present invention will be described with reference to  FIGS. 1 to 8 . In the first embodiment, an ejector  16  of the present invention is typically used for a refrigeration cycle device  10  shown in  FIG. 1 . The refrigeration cycle apparatus  10  shown in  FIG. 1  can be used for a vehicle air conditioner, for example. 
     In the refrigeration cycle device  10 , a compressor  11  is configured to draw refrigerant, to compress the drawn refrigerant, and to discharge the compressed high-pressure and high-temperature refrigerant. The compressor  11  is driven and rotated by a vehicle engine (not shown) via an electromagnetic clutch and a belt, or the like, as an example. 
     The compressor  11  may be a variable displacement compressor in which a discharge capacity of the refrigerant can be continuously adjustable, or may be a fixed displacement compressor in which the discharge capacity of the refrigerant can be adjusted by changing a compressor operation ratio. For example, in the fixed displacement compressor, the compressor operation ratio is changed by interruption of the electromagnetic clutch. Alternatively, an electrical compressor may be used as the compressor  11  such that the refrigerant discharge capacity of the compressor  11  can be adjusted by adjusting a rotation speed of an electrical motor. 
     A refrigerant radiator  12  used as a heat exchanger for heat radiation such as a refrigerant cooler is disposed at a refrigerant discharge side of the compressor  11 . The radiator  12  is configured to perform a heat exchange between the high-pressure refrigerant discharged from the compressor  11  and outside air (i.e., air outside a vehicle compartment) blown by a blower fan  12   a,  thereby cooling the high-pressure refrigerant in the radiator  12 . 
     As the refrigerant used in a refrigerant cycle of the refrigeration cycle apparatus  10 , a Freon-based refrigerant such as HFC134a may be used so that a refrigerant pressure on a high-pressure side in the refrigerant cycle does not excess the critical pressure of the refrigerant. In this case, the radiator  12  is used as a condenser in which the refrigerant is cooled and condensed therein. 
     A receiver  12   b  is located at a refrigerant outlet side of the radiator  12 . The receiver  12   b  is a gas-liquid separator with a vertically elongated tank. The receiver  12   b  is configured to separate the refrigerant flowing therein into gas refrigerant and liquid refrigerant, and to store therein surplus liquid refrigerant in the refrigerant cycle. The receiver  12   b  has a liquid refrigerant outlet at a lower side of the tank so that the liquid refrigerant flows out of the receiver  12  from the liquid refrigerant outlet. As an example of the present embodiment, the receiver  12   b  is formed integrally with the radiator  12 , as shown in  FIG. 1 . However, the receiver  12   b  may be located separately from the radiator  12 , or may be omitted. 
     A branch portion  13  is connected to the liquid refrigerant outlet of the receiver  12   b,  and is configured to divide the liquid refrigerant from the receiver  12   b  into two streams. For example, the branch portion  13  is a three-way joint member having one refrigerant inlet and first and second refrigerant outlets. The three-way joint member used as the branch portion  13  may be configured by bonding pipes having different pipe diameters, or may be configured by providing plural refrigerant passages in a metal block member or a resin block member. 
     One of the refrigerant streams branched at the branch portion  13  flows into a first refrigerant passage  14   a  (nozzle-side refrigerant passage), and the other one of the refrigerant streams branched at the branch portion  13  flows into a second refrigerant passage  14   b  (suction-side refrigerant passage). One end of the first refrigerant passage  14   a  is connected to the first refrigerant outlet of the branch portion  13 , and the other end of the first refrigerant passage  14   a  is connected to an inlet of a nozzle  16   a  of the ejector  16 , so that one of the refrigerant streams branched at the branch portion  13  flows into the nozzle  16   a  through the first refrigerant passage  14   a.  One end of the second refrigerant passage  14   b  is connected to the second refrigerant outlet of the branch portion  13 , and the other end of the second refrigerant passage  14   b  is connected to a refrigerant suction port  16   d  of the ejector  16 , so that the other one of the refrigerant streams branched at the branch portion  13  flows into the refrigerant suction port  16   d  through the second refrigerant passage  14   b.    
     An expansion valve  15  is located in the first refrigerant passage  14   a  at an upstream side of the nozzle  16   a  of the ejector  16  in a refrigerant flow of the first refrigerant passage  14   a.  The expansion valve  15  is used as a decompression portion configured to decompress high-pressure liquid refrigerant flowing from the receiver  12   b  into the first refrigerant passage  14   a  to be in a gas-liquid two-phase state having a middle pressure. The expansion valve  15  is also used as a flow amount adjusting portion for adjusting a flow amount of the refrigerant flowing into the nozzle  16   a.    
     In the present embodiment, the expansion valve  15  is a thermal expansion valve having a temperature sensing portion  15   a  that is located at a refrigerant suction passage of the compressor  11  so as to detect a super-heat degree of the refrigerant to be drawn into a refrigerant suction side of the compressor  11 . In this embodiment, the refrigerant on the refrigerant suction side of the compressor  11  corresponds to the refrigerant at a refrigerant outlet side of a first evaporator  17 . That is, the temperature sensing portion  15   a  detects the super-heat degree of the refrigerant at a refrigerant outlet side of the first evaporator  17  based on at least one of temperature and pressure of the refrigerant at the refrigerant outlet side of the first evaporator  17 , and a valve open degree of the expansion valve  15  is adjusted using a mechanical mechanism or an electrical mechanism so that the super-heat degree of the refrigerant at the refrigerant outlet of the first evaporator  17  is approached to a predetermined value. Thus, the flow amount of the refrigerant flowing to downstream of the expansion valve  15  can be adjusted. 
     The other throttle structure or decompression device may be used instead of the thermal expansion valve  15 . For example, a decompression device such as an electrical variable throttle device or a fixed throttle device, or the other type expansion valve may be used instead of the thermal expansion device  15 . 
     The elector  16  is located at a refrigerant outlet side of the expansion valve  15 . The ejector  14  is adapted as a decompression portion for further decompressing the refrigerant flowing from the expansion valve  15 , and as a refrigerant circulation portion for circulating the refrigerant by the suction action of a high-speed refrigerant flow jetted from the nozzle  16   a.  The structure of the ejector  16  will be now described with reference to  FIGS. 2A to 4 . 
       FIG. 2A  is an axial sectional view of the ejector  16  taken along a section including an axial line,  FIG. 2B  is a cross-sectional view taken along the line IIB-IIB in  FIG. 2A  at an inlet of a suction passage  16 i of the ejector  16 , and  FIG. 2C  is a cross-sectional view taken along the line IIC-IIC in  FIG. 2A  at an outlet of the suction passage  16 i of the ejector  16 . 
     As shown in  FIG. 2A , the ejector  16  is configured by the nozzle  16   a  and a body portion  16   b.  The nozzle  16   a  is made of a metal such as a stainless alloy, and is formed into an approximately cylindrical shape having a taper end portion tapered toward downstream in the refrigerant flow. The refrigerant passage sectional area of the nozzle  16   a  is changed in the refrigerant flow direction so that the refrigerant flowing into the nozzle  16   a  is decompressed and expanded in iso-entropy. 
     A refrigerant jet port  16   c,  from which the refrigerant is jetted from the nozzle  16   a,  is formed at a tip end of the taper end portion of the nozzle  16   a.  The nozzle  16   a  is disposed in the body portion  16   b  and is attached into the body portion  16   b  such that the refrigerant is prevented from being leaked from a fixing portion of the nozzle  16   a  and the body portion  16   b.  For example, the nozzle  16   a  may be air-tightly fitted into the body portion  16   b,  or may be air-tightly bonded to the body portion  16   b  by using a bonding means such as welding, pressing or brazing, or the like. 
     For example, the nozzle  16   a  may be a Laval nozzle having a throat portion at which the passage sectional area becomes smallest within the refrigerant passage inside of the nozzle  16   a.  The nozzle  16   a  is configured such that the flow velocity of the refrigerant jetted from the refrigerant jet port  16   c  of the nozzle  16   a  becomes equal to or larger than the sound velocity. Alternatively, the nozzle  16   a  may be configured by a taper nozzle so that the flow velocity of the refrigerant jetted from the refrigerant jet port  16   c  of the nozzle  16   a  becomes equal to or larger than the sound velocity. 
     The body portion  16   b  can be made of a metal, for example, aluminum or an aluminum alloy, or may be made of a material other than the metal such as resin. The body portion  16   b  is provided with the refrigerant suction port  16   b  penetrating through the interior and the exterior of the body portion  16   b  in a radial direction perpendicular to the axial direction of the nozzle  16   a.  The refrigerant suction port  16   b  is open in the body portion  16   b  at a portion radially outside of the nozzle  16   a.  The body portion  16   b  has therein a mixing and pressurizing portion  16   e  extending in the axial direction (longitudinal direction) from a position of the refrigerant jet port  16   c  to the refrigerant outlet (downstream end). 
     The refrigerant suction port  16   d  is coupled to a refrigerant outlet side of a second evaporator  19  so that the refrigerant from the second evaporator  19  is drawn into the suction passage  16   i  from the refrigerant suction port  16   d.  The refrigerant suction port  16   d  is provided at an outer peripheral side of the nozzle  16   a,  and communicates with a space at the refrigerant jet port  16   c  within the body portion  16   b  through the suction passage  16   i.    
     An inlet space, into which refrigerant from the refrigerant suction port  16   d  flows, is provided within the body portion  16   b  around the refrigerant suction port  16   d,  and the suction passage  16   i  is provided between the outer wall surface of the taper end portion of the nozzle  16   a  and an inner wall surface of the body portion  16   b.  Therefore, the refrigerant flowing from the refrigerant suction port  16   d  into the inlet space of the body portion  16   b  is introduced into an inlet of the mixing and pressurizing portion  16   e  via the suction passage  16   i.  Here, the inlet of the mixing and pressurizing portion  16   e  in the body portion  16   b  substantially corresponds to the position of the refrigerant jet port  16   c  of the nozzle  16   a  in the axial direction. 
       FIG. 2B  shows a refrigerant passage sectional area Ain at the inlet of the suction passage  16   i,  and  FIG. 2C  shows a refrigerant passage sectional area Aout at the outlet of the suction passage  16   i.  As shown in  FIGS. 2B and 2C , refrigerant passage area Aout at the outlet of the suction passage  16   i  is smaller than the refrigerant passage area Ain at the inlet of the suction passage  16   i.    
       FIG. 3  shows a variation in a ratio (passage area ratio) of a refrigerant passage sectional area of the suction passage  16   i  in the refrigerant flow direction, to the refrigerant passage sectional area at the inlet of the suction passage  16   i.  As shown by the solid line in  FIG. 3 , the passage area ratio of the suction passage  16   i  is gradually reduced from the inlet to the outlet of the suction passage  16   i  in the refrigerant flow of the suction passage  16   i.  As shown in  FIG. 3 , a reduce degree of the passage sectional area on a side of the inlet of the suction passage  16   i  is larger than a reduce degree of the passage sectional area on a side of the outlet of the suction passage  16   i.    
     Specifically, as shown by the solid-line graph of  FIG. 3 , the passage sectional area of the suction passage  16   i  is rapidly reduced in a range from the inlet of the suction passage  16   i  approximately to a middle position of the suction passage  16   i,  and the passage sectional area of the suction passage  16   i  is slowly reduced approximately from the middle position of the suction passage  16   i  to the outlet of the suction passage  16   i.  Thus, as compared with the comparative chain line straightly connecting the inlet and the outlet of the suction passage  16   i,  the variation line (i.e., solid line in  FIG. 3 ) of the passage area ratio of the suction passage  16   i  is positioned under the comparative chain line and is made convex downwardly. 
     In the present embodiment, the passage sectional area of the suction passage  16   i  is changed as described above so that the flow velocity of the refrigerant passing through the suction passage  16   i  becomes equal to or greater than the sound velocity. Thus, the flow velocity of the suction refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  can be made approximately equal to the flow velocity of the jet flow jetted from the jet port  16   c  of the nozzle  16   a  into the mixing and pressurizing portion  16   e.  Accordingly, it is possible for the suction refrigerant can be decompressed in iso-entropy in the suction passage  16   i.    
     As shown in  FIG. 2A , the mixing and pressurizing portion  16   e  is positioned just downstream of the nozzle  16   a  and the suction passage  16   i,  so that the kinetic energy of gas-liquid two-phase refrigerant is converted to the pressure energy thereof in the mixing and pressurizing portion  16   e  while the jet refrigerant jetted from the nozzle  16   a  and the suction refrigerant drawn from the refrigerant suction port  16   d  are mixed in the mixing and pressurizing portion  16   e.    
     The mixing and pressurizing portion  16   e  is configured by a straight portion  16   g  from the inlet of the mixing and pressurizing portion  16   e  to a predetermined range, and an expanding portion  16   h  from the downstream side of the straight portion  16   g  to the outlet of the ejector  16 . The straight portion  16   g  of the mixing and pressurizing portion  16   e  is a cylindrical passage having a constant passage sectional area. The expanding portion  16   h  is gradually enlarged in the passage sectional area from its inlet toward downstream. 
     The straight portion  16   g  is provided in a range from the inlet of the mixing and pressurizing portion  16   e,  such that the flow velocity of the gas refrigerant and the flow velocity of the liquid refrigerant in the refrigerant flowing into the mixing and pressurizing portion  16   e  become approximately equal to each other in the straight portion  16   e.  When the length of the straight portion  16   g  in the axial direction of the nozzle  16   a  is L 1  and when the length of the mixing and pressurizing portion  16   e  in the axial direction of the nozzle  16   a  is L 2 , a ratio L 1 /L 2  is set about 0.2, as an example. 
     The refrigerant passage shape of the expanding portion  16   h  in the section including the center line (axial line) is changed in a smoothly curved line as shown in  FIG. 4 . An increase degree in the refrigerant passage sectional area of the expanding portion  16   h  is changed as shown in  FIG. 4 . The increase degree on an inlet side of the refrigerant passage sectional area of the expanding portion  16   h  is larger than the increase degree on an outlet side of the refrigerant passage sectional area of the expanding portion  16   h.  That is, the increase degree on the inlet side of the refrigerant passage sectional area of the expanding portion  16   h  is relatively rapidly increased, and the increase degree on the outlet side of the refrigerant passage sectional area of the expanding portion  16   h  is relatively slowly increased, as compared with the mean increase degree from the inlet to the outlet of the expanding portion  16   h.    
     As shown in  FIG. 4 , the passage wall surface of the expanding portion  16   h  on the section including the axial line of the expanding portion  16   h,  the sectional shape at the inlet side of the passage wall surface of the expanding portion  16   h  is formed into a curved line  101  with a slight convex toward an inner peripheral side, and the sectional shape at the outlet side of the passage wall surface of the expanding portion  16   h  is formed into a curved line  102  with a slight convex toward an outer peripheral side. The straight portion  16   g  and the expanding portion  16   h  of the mixing and pressurizing portion  16   e  are continuously extended, and are configured such that the refrigerant is substantially pressurized in iso-entropy in the mixing and pressurizing portion  16   e  while the refrigerant is prevented from being separated from the passage wall surface of the mixing and pressurizing portion  16   e  at the outlet of the mixing and pressurizing portion  16   e.    
     Thus, the energy loss of the refrigerant while passing through the mixing and pressurizing portion  16   e  can be reduced, and the energy loss of the refrigerant when flowing out of the mixing and pressurizing portion  16   e  can be reduced.  FIG. 4  is a schematic diagram for only explaining the sectional shape of the inner wall surface of the mixing and pressurizing portion  16   e,  and the black points in  FIG. 4  are indicated only for explaining the positions of the straight portion  16   g,  the curved line  101  and the curved line  102  in the section shape of the mixing and pressurizing portion  16   e.    
     As shown in  FIG. 1 , the first evaporator  17  is connected to the downstream side of the mixing and pressurizing portion  16   e  of the ejector  16 , that is, the outlet side of the expanding portion  16   h  of the mixing and pressurizing portion  16   e.  The first evaporator  17  is a heat exchanger, in which the refrigerant flowing out of the mixing and pressurizing portion  16   e  of the ejector  16  is heat-exchanged with air blown by a blower fan  17   a,  and is evaporated by absorbing heat from air passing through the first evaporator  17 . 
     The blower fan  17   a  may be an electrical blower in which a fan rotation speed is controlled by a control voltage output from an air conditioning controller (not shown) so as to control an air blowing amount. A refrigerant outlet of the first evaporator  17  is coupled to the refrigerant suction port of the compressor  11 . 
     In contrast, the second passage  14   b  has the one end branched from the first passage  14   a  at the branch portion  13 , and the other end connected to the refrigerant suction port  16   d  of the ejector  16 . A throttle unit  18  and the second evaporator  19  are located in the second passage  14   b  between the branch portion  13  and the refrigerant suction port  16   d  of the ejector  16 . The throttle unit  18  is configured to function as a decompression means for decompressing the refrigerant flowing into the second evaporator  19  via the second passage  14   b  as well as a flow amount adjusting means for adjusting a flow amount of the refrigerant flowing into the second evaporator  19 . As the throttle unit  18 , a fixed throttle such as a capillary tube an orifice or the like may be used, or a variable throttle may be used. 
     The second evaporator  19  is located in the second passage  14   b  at a downstream side of the throttle unit  18 , so that the refrigerant decompressed in the throttle unit  18  flows into the second evaporator  19 . The second evaporator  19  is a heat exchanger, in which the refrigerant flowing out of the throttle unit  18  is heat-exchanged with air blown by a blower fan  19   a,  and is evaporated by absorbing heat from air passing through the second evaporator  19 . The blower fan  19   a  may be an electrical blower, similarly to the blower fan  17   a.    
       FIG. 5A  shows a Mollier diagram showing refrigerant states in the refrigerant cycle of the refrigeration cycle device  10  with the above structure according to the first embodiment, and  FIG. 5B  is an enlarged view showing the part VB in  FIG. 5A . When the compressor  11  is driven and is operated by a power source such as a vehicle engine, the high-temperature and high-pressure refrigerant is discharged from the compressor  100  (point  201  in  FIG. 5A ), and flows into the radiator  12 . The high-temperature and high-pressure refrigerant is cooled and condensed in the radiator  12  (from point  201  to point  202  in  FIG. 5A ). 
     The high-pressure refrigerant flowing out of the radiator  12  flows into the receiver  12   b,  and is separated into gas refrigerant and liquid refrigerant. The separated liquid refrigerant flowing out of the receiver  12   b  flows into the branch portion  13  (from point  202  to point  203  in  FIG. 5A ), and is branched into a refrigerant stream flowing into the first passage  14   a  so as to flow toward the nozzle  16   a  and a refrigerant stream flowing into the second passage  14   b  so as to flow toward the refrigerant suction port  16   d.    
     A flow amount ratio Ge/Gnoz of the flow amount Ge of the refrigerant flowing through the second passage  14   b  to the flow amount Gnoz of the refrigerant flowing through the first passage  14   a  is determined based on flow characteristics (decompression characteristics) of the expansion valve  15 , the nozzle  16   a  of the ejector  16  and the throttle unit  18 . 
     The refrigerant flowing into the expansion valve  15  through the branched first passage  14   a  is decompressed and expanded in the expansion valve  15  while the flow amount of the refrigerant to flow into the nozzle  16   a  of the ejector  16  is adjusted by the expansion valve  15  (from point  203  to point  204  in  FIG. 5A ). Here, the flow amount of the refrigerant is adjusted by the expansion valve  15  so that the super-heat degree of the refrigerant at the refrigerant outlet side (the point  208  of  FIG. 5A ) of the first evaporator  17  is approached to a predetermined value. As shown in  FIG. 5A  from point  203  to point  204 , the refrigerant is decompressed in iso-enthalpy in the expansion valve  15 . 
     The refrigerant after being decompressed in the expansion valve  15  is further decompressed in the nozzle  16   a  substantially in iso-entropy while the enthalpy of the refrigerant is reduced (from point  204  to point  205  in  FIG. 5A ). The pressure energy of the refrigerant is converted to the speed energy of the refrigerant in the nozzle  16   a  so that the refrigerant is jetted from the refrigerant jet port  16   c  of the nozzle  16   a  by a high speed. 
     By the high-speed refrigerant stream from the refrigerant jet port  16   c  of the nozzle  16   a,  the refrigerant evaporated in the second evaporator  19  is drawn into the ejector  16  from the refrigerant suction port  16   d.  In  FIG. 5A , ΔH indicates a reduction part of the enthalpy while the refrigerant is decompressed and expanded in iso-entropy at the nozzle  16   a.    
     The refrigerant jetted from the nozzle  16   a  and the refrigerant drawn from the refrigerant suction port  16   d  flows into the mixing and pressurizing portion  16   e  positioned downstream of the nozzle  16   a.  The refrigerant jetted from the nozzle  16   a  and the refrigerant drawn from the refrigerant suction port  16   d  are mixed in the mixing and pressurizing portion  16   e,  and the speed energy of the refrigerant is converted to the pressure energy, thereby increasing the refrigerant pressure in the mixing and pressurizing portion  16   e  (point  205 →to point  206 →point  207  in  FIG. 5A ). 
     The refrigerant flowing out of the mixing and pressurizing portion  16   e  of the ejector  16  flows into the first evaporator  17 . In the first evaporator  17 , low-pressure refrigerant is evaporated by absorbing heat from air blown by the blower fan  17   a,  so that the enthalpy of the refrigerant is increased (from point  207  to point  208  in  FIG. 5A ). Thus, air passing through the first evaporator  17  is cooled and the cooled air can be blown into a compartment to be cooled (e.g., a vehicle compartment). The gas refrigerant flowing out of the first evaporator  17  is drawn into the compressor  11  to be compressed again by the compressor  11  (from point  208  to point  201  in  FIG. 5A ). 
     In contrast, the refrigerant stream flowing into the second passage  14   b  from the branch portion  13  is decompressed and expanded by the throttle unit  18  (from point  203  to point  209  in  FIG. 5A ), and low-pressure refrigerant decompressed by the throttle unit  18  flows into the second evaporator  19 . In the second evaporator  19 , low-pressure refrigerant is evaporated by absorbing heat from air blown by the blower fan  19   a,  so that the enthalpy of the refrigerant is increased (from point  209  to point  210  in  FIG. 5A ). Thus, air passing through the second evaporator  19  is cooled and the cooled air can be blown into a compartment to be cooled (e.g., the vehicle compartment). 
     The refrigerant after passing through the second evaporator  19  is drawn into the ejector  16  from the refrigerant suction port  16   d.  The refrigerant drawn from the refrigerant suction port  16   d  flows into the mixing and pressurizing portion  16   e  of the ejector  16  through the suction passage  16   i.  In the present embodiment, the flow velocity of the refrigerant flowing through the suction passage  16   i  is greater than the sound velocity, and the refrigerant passing through the suction passage  16   i  is decompressed in iso-entropy as shown in  FIG. 5B  from point  210  to point  210 ′. While the refrigerant is decompressed in the suction passage  16   i  in iso-entropy, the enthalpy of the refrigerant is reduced by Δh. 
     The refrigerant flowing from the refrigerant suction port  16   d  into the mixing and pressurizing portion  16   e  through the suction passage  16   i  is mixed with the refrigerant jetted from the nozzle  16   a  in the mixing and pressurizing portion  16   e.  (from point  210 ′ to point  206  in  FIG. 5A ). Then, the refrigerant is pressurized in the mixing and pressurizing portion  16   e  (from point  206  to point  207  in  FIG. 5A ), and is supplied to the first evaporator  17  after passing through the mixing and pressurizing portion  16   e.    
     In the refrigeration cycle device  10  having the ejector  16 , the refrigerant flowing out of the mixing and pressurizing portion  16   e  of the ejector  16  can be supplied to the first evaporator  17  while the refrigerant decompressed by the throttle unit  18  in the second passage  14   b  can be supplied to the second evaporator  19  through the throttle unit  18 . Thus, both the first evaporator  17  and the second evaporator  19  can be operated simultaneously to have cooling functions. 
     Because a refrigerant downstream side of the first evaporator  17  is connected to the refrigerant suction side of the compressor  11 , the refrigerant pressurized in the mixing and pressurizing portion  16   e  of the ejector  16  is drawn into the compressor  11 . Therefore, the suction pressure of the compressor  11  can be increased, and the driving power of the compressor  11  can be reduced. As a result, the coefficient of performance (COP) in the refrigerant cycle of the refrigeration cycle device  10  can be effectively improved. 
     In the ejector  16  of the first embodiment, the suction passage  16   i  is provided to decompress the refrigerant in iso-entropy, such that the flow velocity of the refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is equal to or larger than the sound velocity. Therefore, the flow velocity of the refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  can be made substantially equal to the flow velocity of the refrigerant jetted from the refrigerant jet port  16   c  of the nozzle  16   a  into the mixing and pressurizing portion  16   e.  Therefore, the flow velocity of the refrigerant drawn from the refrigerant suction port  16   d  can be increased while the energy loss of the refrigerant passing through the suction passage  16   i  can be reduced. 
     Accordingly, the flow velocity of the gas refrigerant flowing into the straight portion  16   g  of the mixing and pressurizing portion  16   e  can be increased, and thereby the terminal velocity of grains of the liquid refrigerant can be increased. 
     Thus, even when gas-liquid two-phase refrigerant passes through the mixing and pressurizing portion  16   e  in the ejector  16 , the pressurizing amount of gas refrigerant can be increased in the mixing and pressurizing portion  16   e,  thereby improving the ejector efficiency ηe. That is, even in the ejector  16  in which the kinetic energy of gas-liquid two-phase refrigerant is converted to the pressure energy thereof, the pressurizing amount of the gas refrigerant can be effectively increased in the mixing and pressurizing portion  16   e.    
     In the present embodiment, because the refrigerant from the refrigerant suction port  16   d  is decompressed in the suction passage  16   i  in iso-entropy as shown in  FIG. 5B , the energy to be used for pressurizing can be increased by the Δh, as compared with a case where the refrigerant is decompressed in iso-enthalpy. Thus, the pressurizing amount in the mixing and pressurizing portion  16   e  can be increased by an amount corresponding to the Δh. 
     The ejector efficiency ηe′ of the present embodiment can be defined as in the following formula (F2) which is different from the formula (F1). 
       η e ′=(( Gnoz +Ge)×(Δ P/ ρ)/( Gnoz×Δi +Ge×Δ h )   (F2) 
     Here, Ge is the flow amount of the suction refrigerant in the suction passage  16   i,  Gnoz is the flow amount of the jet refrigerant jetted from the nozzle  16   a,  ΔP is the pressurizing amount in the mixing and pressurizing portion  16   e,  ρ is the density of the suction fluid, Δi is the enthalpy difference between the inlet and the outlet of the nozzle  16   a,  and Δh is the energy to be used for pressurizing. As compared with the above formula (F1), the expansion energy item (Ge×Δh) in the suction passage  16   i  can be added in the denominator item (recovery energy item) in the formula (F2). 
     Thus, in the present embodiment, if the various configurations of the ejector  16  are set such that the same ejector efficiency ηe in formula F1 is obtained, the pressurizing amount ΔP can be increased by the recovery energy, thereby effectively improving the ejector efficiency. 
     In the present embodiment, the enthalpy reduction part Δh of the refrigerant while being decompressed and expanded in iso-entropy in the suction passage  16   i,  and the enthalpy reduction part ΔH of the refrigerant while being decompressed and expanded in iso-enthalpy in the nozzle  16   a  have the following relationship in the formula F3. 
       ΔH≧Δh   (F3) 
     That is, in the present embodiment, respective configurations of the ejector  16  are set to satisfy the above formula F 3 . That is, ΔH is the enthalpy difference between the enthalpy of the refrigerant at the inlet of the nozzle  16   a  and the enthalpy of the refrigerant at the refrigerant jet port  16   c  of the nozzle  16   a,  and Δh is the enthalpy difference between the enthalpy of the refrigerant at the inlet of the suction passage  16   i  and the enthalpy of the refrigerant at the outlet of the suction passage  16   i.    
     According to the first embodiment of the present invention, because the refrigerant is decompressed and expanded in the suction passage  16   i  in iso-entropy, the flow velocity of the refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  can be increased. If the iso-entropy decompression amount of the refrigerant in the suction passage  16   i  is increased more than a necessary amount, the flow velocity of the refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  is unnecessarily increased as compared with the flow velocity of the refrigerant jetted from the nozzle  16   a  into the mixing and pressurizing portion  16   e.  Therefore, the energy loss may be increased while the gas refrigerant and the liquid refrigerant having different flow velocities are mixed in the mixing and pressurizing portion  16   e,  and thereby the ejector efficiency may be decreased. 
     That is, the unnecessary increased flow velocity of the refrigerant in the suction passage  16   i  causes the gas refrigerants having different flow speeds to be mixed in the mixing and pressurizing portion  16   e,  thereby increasing the energy loss and decreasing the ejector efficiency. 
       FIG. 6  shows variations in the flow velocities of the gas refrigerant and liquid refrigerants within the ejector  16  when ΔH≧Δh and when ΔH&lt;Δh. In the graphs of  FIG. 6 , the horizontal axis indicates axial positions in the ejector  16  from the inlet of the nozzle  16   a  to the outlet of the ejector  16 . The upper side graph of  FIG. 6  indicates the present embodiment where ΔH≧Δh, in which GJ 2  indicates variations in the flow velocity of gas refrigerant in the jet refrigerant jetted from the nozzle  16   a,  GS 2  indicates variations in the flow velocity of gas refrigerant in the suction refrigerant drawn from the refrigerant suction port  16   d,  the chain line of LIQUID indicates variations in the flow velocity of liquid refrigerant. The lower side graph of  FIG. 6  indicates a comparison example where ΔH&lt;Δh, in which GJ 1  indicates variations in the flow velocity of gas refrigerant in the jet refrigerant jetted from the nozzle  16   a,  GS 1  indicates variations in the flow velocity of gas refrigerant in the suction refrigerant drawn from the refrigerant suction port  16   d,  the chain line of LIQUID indicates variations in the flow velocity of liquid refrigerant. 
     More specifically, the upper side graph in  FIG. 6  shows the first embodiment of the present invention in which the flow velocity GS 2  of the gas refrigerant in the suction refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is approximately equal to the flow velocity GJ 2  of gas refrigerant in the jet refrigerant flowing into the mixing and pressurizing portion  16   e  from the nozzle  16   a,  and thereby ΔH≧Δh. 
     In contrast, the lower side graph in  FIG. 6  shows the comparison example in which the flow velocity GS 1  of the gas refrigerant in the suction refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is greatly faster than the flow velocity GJ 1  of gas refrigerant in the jet refrigerant flowing into the mixing and pressurizing portion  16   e  from the nozzle  16   a,  and thereby ΔH&lt;Δh. 
     As shown in the graphs of  FIG. 6 , if the flow velocity of the gas refrigerant in the suction refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is greatly faster than the flow velocity of gas refrigerant in the jet refrigerant flowing into the mixing and pressurizing portion  16   e  from the nozzle  16 , the flow of the gas refrigerant in the jet refrigerant is accelerated by the flow of the gas refrigerant in the suction refrigerant. When the flow velocity of the gas refrigerant in the jet refrigerant becomes equal to the flow velocity of the gas refrigerant in the suction refrigerant, the gas refrigerant in the jet refrigerant and the gas refrigerant in the suction refrigerant are joined with the same flow velocity. Then, after the gas refrigerant in the jet refrigerant and the gas refrigerant in the suction refrigerant are joined with the same flow velocity, the grains of the liquid refrigerant in the jet refrigerant are accelerated by the joined gas refrigerant. 
     Accordingly, the terminal velocity, at which the flow velocity of the liquid refrigerant reaches to the joined flow velocity of the gas refrigerant in the jet refrigerant and the gas refrigerant in the suction refrigerant, is positioned on a downstream side in the mixing and pressurizing portion  16   e,  and thereby the moving distance of the liquid refrigerant from the inlet to a position corresponding to the terminal velocity in the mixing and pressurizing portion  16   e  is increased. As a result, the distance from the position corresponding to the terminal velocity to the outlet of the mixing and pressurizing portion  16   e,  in which the kinetic energy is transmitted between the gas refrigerant and the liquid refrigerant after the grains of the liquid refrigerant reaches to the terminal velocity in the mixing and pressurizing portion  16   e,  becomes shorter, and thereby the refrigerant can not be sufficiently pressurized in the mixing and pressurizing portion  16   e.    
     In contrast, according to the first embodiment of the present invention, the configurations including dimensions in the respective portions of the ejector  16  are set so that ΔH≧Δh. Thus, it can prevent the flow velocity of the suction refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  from being excessively increased. 
     More specifically, the tilt of the iso-entropy line of the gas refrigerant from the inlet to the outlet of the suction passage  16   i  (point  210  to point  210 ′ of  FIGS. 5A ,  5 B) relative to a horizontal line is smaller than the tilt of the iso-entropy line of the gas refrigerant from the inlet to the outlet of the nozzle (point  204  to point  205  of  FIG. 5A ) relative to the horizontal line. Therefore, the decompression amount of the refrigerant in the suction passage  16   i  can be accurately set smaller than the decompression amount of the refrigerant in the nozzle  16   a.  Thus, the refrigerant can be decompressed in the suction passage  16   i  by a suitable decompression amount. 
     As a result, the energy loss, generated while gas refrigerants having different flow velocities are mixed, can be reduced, and the refrigerant can be sufficiently pressurized in the mixing and pressurizing portion  16   e,  thereby effectively improving the ejector efficiency. 
     According to the first embodiment of the present invention, because the straight portion  16   g  is provided in a suitable range at a refrigerant inlet side of the mixing and pressurizing portion  16   e,  the energy force of the gas refrigerant can be effectively applied to the grains of the liquid refrigerant in the straight portion  16   g,  and thereby the flow velocity of the grains of the liquid refrigerant can rapidly reach to the terminal velocity in the straight portion  16   g.    
     Furthermore, the kinetic energy of the liquid refrigerant having being reached to the terminal velocity can be effectively transmitted to the gas refrigerant in the expanding portion  16   h.  As a result, the energy transmission loss between the gas refrigerant and the liquid refrigerant in the expanding portion  16   h  can be reduced, and thereby the ejector efficiency can be sufficiently improved. 
       FIG. 7A  shows the flow velocity of the gas refrigerant, the flow velocity of the liquid refrigerant, the pressurizing amount ΔP shown by the line P 1 , at respective positions from the inlet to the outlet of the mixing and pressurizing portion  16   e  of the ejector  16  according to the present embodiment. On the other hand,  FIG. 7B  shows the flow velocity of the gas refrigerant, the flow velocity of the liquid refrigerant, the pressurizing amount ΔP shown by the line P 2 , at respective positions from the inlet of a mixing portion to the outlet of a diffuser portion of an ejector of a comparative example. 
     As shown in  FIG. 7A , the straight portion  16   g  is provided in a range of the mixing and pressurizing portion  16   e  from the inlet of the mixing and pressurizing portion  16   e,  such that the flow velocities of the gas refrigerant and the liquid refrigerant in the refrigerant flowing into the mixing and pressurizing portion  16   e  become equal at the downstream end of the straight portion  16   g.  That is, the terminal velocity is caused at the downstream end of the straight portion  16   g.  Therefore, the kinetic energy of the refrigerant immediately after reaching to the terminal velocity can be converted to the pressure energy in the expanding portion  16   h.    
     Because the flow velocity of the liquid refrigerant reaches to the terminal velocity at the inlet side of the expanding portion  16   h,  the energy transmission loss between the gas refrigerant and the liquid refrigerant can be effectively reduced. Thus, the flow velocity of the liquid refrigerant and the gas refrigerant at the outlet of the expanding portion  16   h  can be sufficiently reduced, and the ratio of energy to be used actually for pressurizing can be increased. 
     As a result, the pressurizing amount ΔP of the refrigerant in the mixing and pressurizing portion  16   e  can be increased in the present embodiment shown by the graph P 1  in  FIG. 7A , as compared with the comparative example shown by the graph P 2  in  FIGS. 7A and 7B . 
     According to experiments by the inventors of the present application, when the ratio L 1 /L 2  of the length L 1  of the straight portion  16   g  to the length L 2  of the mixing and pressurizing portion  16   e  from the inlet to the outlet of the mixing and pressurizing portion  16   e  is set about at 0.2, the pressurizing amount ΔP of the refrigerant in the mixing and pressurizing portion  16   e  can be made maximum. 
     When the ratio L 1 /L 2  is set about at 0.2, the flow velocities of the gas refrigerant and the liquid refrigerant flowing out of the outlet of the straight portion  16   g  can be made approximately equal, and the pressurizing amount ΔP of the refrigerant in the mixing and pressurizing portion  16   e  can be made maximum. If the manufacturing error of the ejector  16  and the variation in the flow amount of the refrigerant circulating in the refrigerant cycle of the refrigeration cycle device  10  are considered, the ejector efficiency can be sufficiently increased when 0&lt;L 1 /L 2 ≦0.4. More preferably, the ratio L 1 /L 2  is set such that 0.1&lt;L 1 /L 2 ≦0.3. 
     In a case where 0&lt;L 1 /L 2 ≦0.4, the ejector efficiency can be sufficiently improved even when the gas-liquid density difference of the gas-liquid two-phase refrigerant passing through the mixing and pressurizing portion  16   e  is changed in a wider range of 0.9-600 kg/M 3 . 
     In the first embodiment of the present invention, the refrigerant can be pressurized substantially in iso-entropy in the entire range of the mixing and pressurizing portion  16   e,  and the sectional shape of the mixing and pressurizing portion  16   e  is changed so as to reduce a separation from of the refrigerant at the outlet of the mixing and pressurizing portion  16   e.  Therefore, the energy loss of the refrigerant passing through the mixing and pressurizing portion  16   e  can be reduced, thereby reducing energy loss of the refrigerant flowing out of the mixing and pressurizing portion  16   e.    
     As a result, the ratio of the energy to be actually used for pressuring can be increased among the recovery energy in the ejector  16 .  FIG. 8  shows energy distribution in the mixing and pressurizing portion  16   e  of the ejector  16  according to the first embodiment and the comparison example. In  FIG. 8 , E 1  indicates the energy to be used for pressurizing, E 2  indicates the remaining energy of the refrigerant, E 3  indicates the energy transmission loss, and E 4  indicates the other loss. As shown in  FIG. 8 , according to the first embodiment, the energy to be used for pressurizing in the mixing and pressurizing portion  16   e  can be greatly increased as compared with the comparison example. 
     Second Embodiment 
     A second embodiment of the present invention will be described with reference to  FIG. 9 .  FIG. 9  is a diagram corresponding to  FIG. 3  of the above-described first embodiment. In the second embodiment, the passage sectional area of the suction passage  16   i  is changed such that the ratio (passage area ratio) of the passage sectional area of the suction passage  16   i  to the passage sectional area at the inlet of the suction passage  16   i  is changed as in the straight line graph shown in  FIG. 9 . As shown in  FIG. 9 , the passage sectional area of the suction passage  16   i  is changed from the inlet to the outlet of the suction passage  16   i  by a constant degree. In the second embodiment, the other parts of the ejector  16  are similar to those in the ejector  16  of the above-described first embodiment. 
     According to the second embodiment of the present invention, the suction passage  16   i  of the ejector  16  can be configured, such that the flow velocity of the suction refrigerant flowing from the suction passage  16   i  into the straight portion  16   g  of the mixing and pressurizing portion  16   e  becomes equal to or greater than the sound velocity and the suction refrigerant is decompressed in iso-entropy. Thus, the terminal velocity of the grains of the liquid refrigerant flowing into the straight portion  16   g  in the mixing and pressurizing portion  16   e  can be increased, thereby improving the ejector efficiency. In the second embodiment, the other parts of the ejector  16  are similar to those in the ejector  16  of the above-described first embodiment. 
     Third Embodiment 
     A third embodiment of the present invention will be described with reference to  FIG. 10 .  FIG. 10  is a diagram corresponding to  FIG. 3  of the above-described first embodiment. In the third embodiment, the passage sectional area of the suction passage  16   i  is changed, such that the passage sectional area at the inlet side of the suction passage  16   i  is gradually reduced toward downstream in the refrigerant flow direction from the inlet of the suction passage  16   i,  and the passage sectional area at the outlet side of the suction passage  16   i  is gradually increased toward downstream in the refrigerant flow direction. That is, at a predetermined portion between the inlet and the outlet of the suction passage  16   i,  the passage sectional area of the suction passage  16   i  becomes smallest, as shown in  FIG. 10 . A reduction ratio of the passage sectional area at the inlet side of the suction passage  16   i  is larger than an increase ratio of the passage sectional area at the outlet side of the suction passage  16   i.  At the outlet side of the suction passage  16   i,  the passage sectional area of the suction passage  16   i  is gradually increased, but is not increased more than the passage sectional area at the inlet of the suction passage  16   i.    
     In the third embodiment, the other parts of the ejector  16  are similar to those in the ejector  16  of the above-described first embodiment. 
     According to the third embodiment of the present invention, the suction passage  16   i  of the ejector  16  is configured such that the flow velocity of the suction refrigerant flowing through the suction passage  16   i  becomes equal to or greater than the sound velocity at a contraction position where the refrigerant passage area becomes smallest in the suction passage  16   i.  Thus, the flow velocity of the suction refrigerant can be increased downstream of the contraction position in the suction passage  16   i.  Therefore, the terminal velocity of the grains of the liquid refrigerant flowing into the straight portion  16   g  in the mixing and pressurizing portion  16   e  can be increased, thereby improving the ejector efficiency. 
     Fourth Embodiment 
     A fourth embodiment of the present invention will be described with reference to  FIG. 11 .  FIG. 11  is a schematic diagram corresponding to  FIG. 4  of the above-described first embodiment, showing a refrigerant passage sectional shape of the expending portion  16   h  in a section including the center axis of the nozzle  16   a  of the ejector  16 . As shown in  FIG. 11 , the passage wall surface of the expending portion  16   h  is configured by combining plural straight line portions  103 ,  104 ,  105 ,  106 ,  107 . That is, the expanding portion  16   h  is formed by plural cylindrical passage portions ( 103  to  107 ) each of which has a taper surface. The taper surfaces of the plural cylindrical passage portions ( 103  to  107 ) are suitably combined so as to form the expending portion  16   h  in the mixing and pressurizing portion  16   e.    
     In the fourth embodiment, the other parts of the ejector  16  are similar to those in the ejector  16  of the above-described first embodiment. 
     In the structure of the expanding portion  16   h  according to the fourth embodiment, the energy transmission loss between the gas refrigerant and the liquid refrigerant can be reduced, thereby sufficiently improving the ejector efficiency. In the above example of the fourth embodiment, the structure of the expanding portion  16   h  is used for the ejector  16  according to the first embodiment. However, the structure of the expanding portion  16   h  of the fourth embodiment can be used for the ejector  16  according to any one of the second and third embodiments of the present invention. 
     Fifth Embodiment 
     A fifth embodiment of the present invention will be described with reference to  FIG. 12 .  FIG. 12  is a schematic diagram corresponding to  FIG. 4  of the above-described first embodiment, showing a refrigerant passage sectional shape of the expending portion  16   h  in a section including the center axis of the nozzle  16   a  of the ejector  16 . As shown in  FIG. 12 , the passage wall surface of the expending portion  16   h  is configured by combining plural straight line portions  103 ,  104 ,  105 , and the curved line portion  102 . That is, the expanding portion  16   h  is formed, by plural cylindrical passage portions ( 103  to  105 ) each of which has a taper surface and by the cylindrical passage portion ( 102 ) having a curved surface ( 102 ). 
     In the fifth embodiment, the other parts of the ejector  16  are similar to those in the ejector  16  of the above-described first embodiment. 
     In the structure of the expanding portion  16   h  according to the fifth embodiment, the energy transmission loss between the gas refrigerant and the liquid refrigerant can be reduced, thereby sufficiently improving the ejector efficiency. In the above example of the fifth embodiment, the structure of the expanding portion  16   h  is used for the ejector  16  according to the first embodiment. However, the structure of the expanding portion  16   h  of the fifth embodiment can be used for the ejector  16  according to any one of the second and third embodiments of the present invention. 
     Sixth to Eighth Embodiments 
     A sixth embodiment of the present invention will be described with reference to  FIG. 13 .  FIG. 13  is a schematic diagram corresponding to  FIG. 4  of the above-described first embodiment, showing a refrigerant passage sectional shape of the expending portion  16   h  in a section including the center axis of the nozzle  16   a  of the ejector  16 . As shown in  FIG. 13 , the passage wall surface of the expending portion  16   h  is configured by a single straight line portion  108  having a constant taper angle. That is, the refrigerant passage sectional area of the expanding portion  16   h  in the mixing and pressurizing portion  16   e  is gradually increased toward downstream by a constant expanding degree in the entire length of the expanding portion  16   h.    
     A seventh embodiment of the present invention will be described with reference to  FIG. 14 .  FIG. 14  is a schematic diagram corresponding to  FIG. 4  of the above-described first embodiment, showing a refrigerant passage sectional shape of the expending portion  16   h  in a section including the center axis of the nozzle  16   a  of the ejector  16 . As shown in  FIG. 14 , the passage wall surface of the expending portion  16   h  is configured by combining plural straight line portions  103 ,  104 ,  105 ,  106 ,  109 . The plural straight line portions  103 ,  104 ,  105 ,  106 ,  109  are suitably combined to configure the expanding portion  16   h  such that the expanding degree of the refrigerant passage sectional area of the expanding portion  16   h  is gradually increased. 
     An eighth embodiment of the present invention will be described with reference to  FIG. 15 .  FIG. 15  is a schematic diagram corresponding to  FIG. 4  of the above-described first embodiment, showing a refrigerant passage sectional shape of the expending portion  16   h  in a section including the center axis of the nozzle  16   a  of the ejector  16 . As shown in  FIG. 15 , the passage wall surface of the expending portion  16   h  is configured by a single curved line portion  110  in which the expending angle is gradually increased as toward downstream. 
     In the sixth to eighth embodiments of the present invention, the other parts of the ejector  16  can be made similar to those of the ejector  16  according to the first embodiment, and the ejector efficiency can be increased. In the above examples of the sixth to eighth embodiments of the present invention, the structure of the expanding portion  16   h  is used for the ejector  16  according to the first embodiment. However, the structure of the expanding portion  16   h  according to any one of the sixth to eighth embodiments can be used for the ejector  16  of any one of the second and third embodiments of the present invention. That is, the expanding portion  16   h  of the mixing and pressurizing portion  16   e  according to any one of the fourth to eighth embodiments can be suitably used for the ejector  16  according to any one of the first to third embodiments. 
     Ninth Embodiment 
     In the above-described embodiments of the present invention, the ejector  16  includes the mixing and pressurizing portion  16   e  that is configured by the straight portion  16   g  and the expanding portion  16   h.  However, in a ninth embodiment of the present invention, an ejector  16  is configured without using the straight portion  16   g,  so that the mixing and pressurizing portion  16   e  is configured only by the expanding portion  16   h  as shown in  FIG. 16 . 
       FIG. 16  is an axial sectional view of the ejector  16  according to the ninth embodiment of the present invention, which corresponds to  FIG. 2A  of the first embodiment. That is, the inlet of the expanding portion  16   h  is located at the position corresponding to the refrigerant jet port  16   c  of the nozzle  16   a  in the axial direction of the nozzle  16   a.  The passage wall surface of the expanding portion  16   h  of the ejector  16  shown in  FIG. 16  has the same passage sectional shape of the expending portion  16   h  in the mixing and pressurizing portion  16   e  of the above-described first embodiment shown in  FIG. 4 . Thus, the inner peripheral surface of the expending portion  16   h  is curved to be convex toward radially inside at the inlet side of the expanding portion  16   h,  and is curved to be convex toward radially outside at the outlet side of the expanding portion  16   h.    
     Thus, even when the straight portion  16   g  is omitted in the mixing and pressurizing portion  16   e,  it is possible for the inlet side of the mixing and pressurizing portion  16   e  to have the same function as the straight portion  16   g,  thereby improving the ejector efficiency. The mixing and pressurizing portion  16   e  configured by only the expanding portion  16   h  according to the ninth embodiment can be used for the ejector  16  according to the second or the third embodiment. 
     Furthermore, in a case where the ejector efficiency can be sufficiently increased by increasing the flow velocity of the suction refrigerant in the suction passage  16   i,  the straight portion  16   g  described in any one of the above embodiments may be omitted from the mixing and pressurizing portion  16   e.    
     For example, in the above-described fourth to eighth embodiments shown in  FIGS. 11 to 15 , the mixing and pressurizing portion  16   e  may be configured by only the expending portion  16   h  without using the straight portion  16   g.  In this case, the inlet of the expanding portion  16   h  is located at a position corresponding to the refrigerant jet port  16   c  of the nozzle  16   a  in the ejector  16  according to any one of the fourth to eighth embodiments. 
     Tenth Embodiment 
     In the ejector  16  according to any one of the above-described embodiments, the suction passage  16   i  is provided between the outer peripheral surface of the tip end portion of the nozzle  16   a  and the inner peripheral surface of body portion  16   b.  In the ejector  16  of the tenth embodiment, the nozzle  16   a  is used as a first nozzle  16   a,  and a second nozzle  16   j  is provided for forming a suction passage  16   i  through which the refrigerant drawn from a refrigerant suction port  16   d  flows into the mixing and pressurizing portion  16   e,  as shown in  FIG. 17 . That is, the suction passage  16   i  is defined by the second nozzle  16   j  and the refrigerant suction port  16   d  is provided at the inlet of the second nozzle  16   j,  so that the refrigerant drawn from the suction port  16   d  flows into the mixing and pressurizing portion  16   e  through the suction passage  16   i.    
     As an example of the second nozzle  16   j  in the tenth embodiment, a Laval nozzle may be used. The refrigerant passage sectional area of the suction passage  16   i  of the second nozzle  16   j  can be changed similar to that of the suction passage  16   i  of the third embodiment. In this case, the advantages of the suction passage  16   i  described in the third embodiment can be obtained. 
     Alternatively, the second nozzle  16   j  can be configured by a taper nozzle such that the refrigerant passage sectional area of the suction passage  16   i  of the second nozzle  16   j  is changed similar to that of the suction passage  16   i  of the first or second embodiment described above. In this case, the advantages of the suction passage  16   i  described in the first or second embodiment can be obtained. 
     Eleventh Embodiment 
     In the above-described embodiments, the ejector  16  is typically used for the refrigeration cycle device  10  that is provided with the radiator  12  and the receiver  12   b,  for example, as shown in  FIG. 1 . In the refrigeration cycle device  10 , the radiator  12  provided with the receiver  12   b  is an example of a super-cooled type condenser in which the refrigerant is cooled and condensed. 
     In the eleventh embodiment, the ejector  16  according to any one of the above-described embodiments is used for a refrigeration cycle device having a super-cooled type condenser that is configured by a condensation heat exchanging portion, a receiver portion and a super-cooling heat exchanging portion. Here, the condensation heat exchanging portion is configured to cool and condense the high-pressure refrigerant from the compressor  11 , the receiver portion is configured to separate the refrigerant flowing from the condensation heat exchanging portion into gas refrigerant and liquid refrigerant, and the super-cooling heat exchanging portion is configured to super-cool the saturated liquid refrigerant from the receiver portion. Even in this case, the liquid refrigerant super-cooled in the super-cooling heat exchanging portion can be introduced into the branch portion  13  to be branched at the branch portion  13 . The other parts of the refrigerant cycle structure in the refrigeration cycle device of the eleventh embodiment may be similar to those of the refrigeration cycle device  10  shown in  FIG. 1 . 
       FIG. 18  is a Mollier diagram showing refrigerant states in a refrigerant cycle of the refrigeration cycle device according to the eleventh embodiment, in which the super-cooled type condenser configured by the condensation heat exchanging portion, the receiver portion and the super-cooling heat exchanging portion is used instead of the receiver  12  provided with the receiver  12   b.  In this case, as shown in  FIG. 18 , the liquid refrigerant of a super-cooled state (point  203 ′ of  FIG. 18 ) is branched at the branch portion  13 . 
     Thus, the refrigerant state flowing from the expansion valve  15  into the nozzle  16   a  of the ejector  16  may become in a gas-liquid two-phase state (point  204  of  FIG. 18 ) or in a liquid state (point  204 ′ of  FIG. 18 ). In  FIG. 18 , the parts corresponding to or similar to those in  FIG. 5A  are indicated by the same reference numbers, and the detail explanation thereof is omitted. 
     Even in the refrigeration cycle device with the Mollier diagram shown in  FIG. 18 , the ejector  16  is configured such that, the flow velocity of the suction refrigerant passing through the suction passage  16   i  of the ejector  16  is increased, the flow velocity of the grains of the liquid refrigerant can be rapidly reached to the terminal velocity by the straight portion  16   g  of the mixing and pressurizing portion  16   e,  and the flow velocity of the refrigerant can be sufficiently reduced in the expanding portion  16   h.  Thus, the ejector efficiency can be improved. 
     Thus, even in a case where gas-liquid two-phase refrigerant flows into the nozzle  16   a  or only the liquid refrigerant flows into the nozzle  16   a  in the refrigerant cycle, when the mixed refrigerant, in which the jet refrigerant jetted from the nozzle  16   a  and the suction refrigerant flowing from the suction passage  16   i  are mixed, is in the gas-liquid two-phase state in the ejector  16 , the ejector efficiency can be effectively improved. 
     Twelfth Embodiment 
     In the above-described embodiments, the ejector  16  is used for the refrigerant cycle in which the expansion valve  15  is provided in the first passage  14   a  at an upstream side of the nozzle  16   a  of the ejector  16 , for example, as shown in  FIG. 1 . However, in the twelfth embodiment, the ejector  16  is used for a refrigerant cycle of a refrigeration cycle device in which the expansion valve  15  is omitted from the refrigeration cycle device  10  shown in  FIG. 1 . The other parts of the refrigeration cycle device of the twelfth embodiment are similar to those of the refrigeration cycle device  10  shown in  FIG. 1 . The refrigerant state of the refrigerant cycle is changed as in the Mollier diagram of  FIG. 19  when the refrigeration cycle device according to the twelfth embodiment is operated. 
     Because the expansion valve  15  is not provided in the refrigeration cycle device shown in  FIG. 1 , the refrigerant branched at the branch portion  13  flows into the nozzle  16   a  of the ejector  16  through the first passage  14   a,  and is decompressed and expanded substantially in iso-entropy in the nozzle  16   a  (from point  203  to point  205  of  FIG. 19 ). Even when the ejector  16  of the present invention is used for the refrigeration cycle device in which the refrigerant from the branch portion  13  without being decompressed is firstly decompressed and expanded in the nozzle  16   a  of the ejector  16 , the ejector efficiency can be improved similarly to that in the above-described first embodiment. 
     Alternatively, both the receiver  12   b  and the expansion valve  15  can be omitted from the refrigeration cycle device  10  shown in  FIG. 1 . In this case, gas-liquid two-phase refrigerant flowing out of the radiator  12  is directly branched at the branch portion  13  (point  202  of  FIG. 19 ), and flows into the nozzle  16   a  of the ejector  16  through the first passage  14   a  to be decompressed and expanded substantially in iso-entropy in the nozzle  16   a  of the ejector  16 . Alternatively, a super-cooled type condenser can be used as the radiator  12  similarly to the above described in the eleventh embodiment while the expansion valve  15  is omitted in the refrigeration cycle device  10  shown in  FIG. 1 . In this case, a super-cooled liquid refrigerant (point  203 ′ of  FIG. 19 ) flowing out of the radiator  12  is branched at the branch portion  13 , and a part of the branched refrigerant flows into the nozzle  16   a  of the ejector  16  through the first passage  14   a  to be decompressed substantially in iso-entropy in the nozzle  16   a.    
     Thirteenth Embodiment 
     In the above-described embodiments, the ejector  16  is used for a refrigerant cycle in which the refrigerant state flowing from the branch portion  13  into the first passage  14   a  and the refrigerant state flowing from the branch portion  13  into the second passage  14   b  are made equal. However, the ejector  16  can be used for a refrigerant cycle in which the refrigerant state flowing from the branch portion  13  into the first passage  14   a  and the refrigerant state flowing from the branch portion  13  into the second passage  14   b  are made different from each other. 
     As an example of a refrigeration cycle device of a thirteenth embodiment, the receiver  12   b  shown in  FIG. 1  is omitted, the expansion valve  15  is located upstream of the branch portion  13 , and the branch portion  13  is configured so as to change the refrigerant states (e.g., dryness) flowing into the first and second passages  14   a,    14   b.    
     For example, the branch portion  13  may be configured to have an interior space in which a scroll flow of the refrigerant is generated so that the dryness distributions of the refrigerant are caused in the interior space of the branch portion  13  by centrifugal force due to the scroll flow of the refrigerant. 
     The first passage  14   a  and the second passage  14   b  are connected to the branch portion  13  so that refrigerant having a predetermined dryness can be respectively introduced into the first passage  14   a  and the second passage  14   b.  Thus, the dryness of the refrigerant flowing into the first passage  14   a  from the branch portion  13  and the dryness of the refrigerant flowing into the second passage  14   b  from the branch portion  13  can be suitably changed. As the structure of the branch portion  13 , the structure described in US 2007/028630 (corresponding to JP 2007-46806) can be incorporated herein by reference. 
     When the refrigeration cycle device according to the thirteenth embodiment is operated, refrigerant states circulated in the refrigerant cycle can be set to be changed as in the Mollier diagram shown in  FIG. 20  or  FIG. 21 . In the diagram of  FIG. 20 , the refrigerant flowing into the nozzle  16   a  of the ejector  16  from the branch portion  13  through the first passage  14   a  is in a gas-liquid two-phase state (point  203 ″ in  FIG. 20 ). On the other hand, in the diagram of  FIG. 21 , the refrigerant flowing into the nozzle  16   a  of the ejector  16  from the branch portion  13  through the first passage  14   a  is in a liquid state (point  203 ′ in  FIG. 21 ). 
     Even in the refrigerant cycle with the operation states shown in  FIG. 20  or  FIG. 21 , the ejector efficiency can be effectively improved by using the ejector  16  according to any one of the first to tenth embodiments. 
     Fourteenth Embodiment 
     In the above-described embodiments, the ejector  16  of the present invention is used for a sub-critical refrigerant cycle in which the pressure of refrigerant on a high-pressure side before being decompressed is lower than the critical pressure of the refrigerant. However, in a fourteenth embodiment of the present invention, the ejector  16  is used for a super-critical refrigerant cycle in which the pressure of refrigerant on the high-pressure side before being decompressed is higher than the critical pressure of the refrigerant. For example, carbon dioxide is used as the refrigerant so that the refrigerant pressure discharged from the compressor  11  becomes higher than the critical pressure of the refrigerant. 
       FIG. 22  shows an example of a refrigeration cycle device  10  according to the fourteenth embodiment of the present invention. In the refrigeration cycle device  10  of  FIG. 22 , the receiver  12   b  and the expansion valve  15  are omitted from the refrigeration cycle device  10  shown in  FIG. 1 , and a pressure control valve is used as the throttle unit  18  as compared with the refrigerant cycle device  10  shown in  FIG. 1 . A valve open degree of the throttle unit  18  is adjusted such that the refrigerant pressure on the high-pressure side of the refrigerant cycle of the refrigeration cycle device  10  is approached to a target pressure that is determined in accordance with a temperature of the refrigerant at a refrigerant outlet side of the radiator  12 . 
     For example, the throttle unit  18  is provided with a temperature sensing portion  18   a  located at the refrigerant outlet side of the radiator  12 . The temperature sensing portion  18   a  is configured to generate therein an inner pressure corresponding to the temperature of the refrigerant on the refrigerant outlet side of the radiator  12 , so that the valve open degree of the throttle unit  18  is adjusted by a balance between the inner pressure of the temperature sensing portion  18   a  and the pressure of the refrigerant on the refrigerant outlet side of the radiator  12 . Thus, the refrigerant pressure on the high-pressure side of the refrigerant cycle can be adjusted to the target pressure, and thereby the COP of the refrigerant cycle can be made maximum. 
     In the fourteenth embodiment, as shown in  FIG. 22 , an accumulator  20  as a low-pressure side gas-liquid separator is located at a refrigerant outlet side of the first evaporator  17  so that surplus refrigerant in the refrigerant cycle is stored in the accumulator  20 . A gas refrigerant outlet is provided in the accumulator  20  and is coupled to the refrigerant suction side of the compressor  11  so that the gas refrigerant separated from the liquid refrigerant in the accumulator  20  is supplied to the compressor  11 . In the components of the refrigeration cycle device  10  shown in  FIG. 22 , the other parts are similar to those of the refrigeration cycle device  10  shown in  FIG. 1 . 
     When the refrigeration cycle device  10  of the present embodiment is operated, the refrigerant state is charged as in the Mollier diagram shown in  FIG. 23 . As shown in  FIG. 23 , the refrigerant is compressed in the compressor  11  to have a pressure higher than the critical pressure of the refrigerant (point  201  of  FIG. 23 ), and is discharged to the radiator  12 . 
     The refrigerant is cooled in the radiator  12  by performing heat exchange with outside air while keeping the refrigerant pressure at the pressure higher than the critical pressure (from point  201  to point  202  of  FIG. 23 ). The high-pressure refrigerant flowing out of the radiator  12  is branched at the branch portion  13  into a refrigerant stream flowing into the first passage  14   a  and a refrigerant stream flowing into the second passage  14   b.    
     The refrigerant flowing into the first passage  14   a  from the branch passage  13  flows through the nozzle  16   a,  the first evaporator  17  and the accumulator  20  in this order (point  202 →point  205 →point  206 →point  207 →point  208  in  FIG. 23 ). The gas refrigerant separated at the accumulator  20  is drawn into the compressor  11 . 
     On the other hand, the refrigerant flowing into the second passage  14   b  flows through the throttle unit  18  (i.e., high-pressure control valve) and the second evaporator  19  in this order, and is drawn into the ejector  16  from the refrigerant suction port  16   d  (point  202 →point  209 →point  210 →point  210 ′→point  206  in  FIG. 23 ). The throttle unit  18  is adjusted so as to adjust the refrigerant pressure on the high pressure side from the refrigerant discharge side of the compressor  11  to the inlet of the nozzle  16   a  of the ejector  16  and the inlet of the throttle unit  18 , such that the COP of the refrigerant cycle becomes the target pressure. 
     Thus, even in the refrigerant cycle of the refrigeration cycle device  10  in which the super-critical refrigerant flows into the nozzle  16   a  of the ejector  16 , the ejector efficiency can be improved. 
     Even in a case where the super-critical refrigerant flows into the nozzle  16   a  of the ejector  16 , when the mixed refrigerant, in which the jet refrigerant jetted from the nozzle  16   a  and the suction refrigerant drawn from the refrigerant suction port  16   d  are mixed, is in a gas-liquid two-phase state in the ejector  16 , the ejector efficiency can be significantly improved. 
     That is, when the ejector  16  is used for a super-critical refrigerant cycle in which at least the jet refrigerant jetted from nozzle  16   d  is in a gas-liquid two-phase state or the refrigerant downstream of the throat portion of the nozzle  16   d  is in a gas-liquid two-phase state, the ejector efficiency can be more significantly improved. 
     Fifteenth Embodiment 
     A fifteenth embodiment of the present invention will be described with reference to  FIGS. 24 and 25 . As shown in  FIG. 24 , in a refrigeration cycle device  10  of the fifteenth embodiment, the compressor  11  is used as a first compressor  11 , and a second compressor  21  is added in the second passage  14   b  between the refrigerant outlet of the second evaporator  19  and the refrigerant suction port  16   d  of the ejector  16 . Therefore, the second compressor  21  compresses the refrigerant flowing out of the second evaporator  19  and discharges the compressed refrigerant to the refrigerant suction port  16   d  of the ejector  16 . The other components of the refrigeration cycle device  10  of the fifteenth embodiment are similar to those of the refrigeration cycle device  10  shown in  FIG. 1 . 
     For example, in the fifteenth embodiment of the present invention, the first evaporator  17  can be used for cooling the interior of a passenger compartment of a vehicle, and the second evaporator  19  can be used for cooling a cooler box (refrigerator) mounted in the vehicle. That is, the space to be cooled by the first evaporator  17  is the passenger compartment of the vehicle, and the space to be cooled by the second evaporator  19  is the interior space of the cooler box. 
     The basic structure of the second compressor  21  may be similar to that of the first compressor  11 , and a generally known compressor may be used as the second compressor  21 . 
       FIG. 25  is a Mollier diagram showing the refrigerant operation state of the refrigerant cycle of the refrigeration cycle device  10 , according to the fifteenth embodiment. As shown in  FIG. 25 , the refrigerant is compressed in the first compressor  11  to be in a high-pressure and high-temperature state (point  201  of  FIG. 25 ), and is discharged to the radiator  12 . The high-pressure and high-temperature refrigerant is cooled in the radiator  12  by performing heat exchange with outside air (from point  201  to point  202  of  FIG. 25 ). The high-pressure refrigerant flowing out of the radiator  12  is separated into gas refrigerant and liquid refrigerant in the receiver  12   b,  and the separated liquid refrigerant flows into the branch portion  13  (from point  202  to point  203  of  FIG. 25 ), similarly to  FIG. 5A  of the first embodiment. Then, the refrigerant is branched at the branch portion  13  into a refrigerant stream flowing into the first passage  14   a  and a refrigerant stream flowing into the second passage  14   b.    
     The refrigerant flowing into the expansion valve  15  through the branched first passage  14   a  is decompressed and expanded in iso-enthalpy by the expansion valve  15  (from point  203  to point  204  in  FIG. 25 ). Then, the refrigerant after being decompressed at the expansion valve  15  is further decompressed and expanded in the nozzle  16   a  substantially in iso-entropy while the enthalpy of the refrigerant is reduced (from point  204  to point  205  in  FIG. 25 ). The pressure energy of the refrigerant is converted to the speed energy of the refrigerant in the nozzle  16   a  so that the refrigerant is jetted from the refrigerant jet port  16   c  by a high speed. Then, the refrigerant jetted from the refrigerant jet port  16   c  of the nozzle  16  is mixed in the mixing and pressurizing portion  16   e  with the refrigerant drawn from the refrigerant suction port  16   d,  so that the mixed refrigerant is pressurized in the mixing and pressurizing portion  16   e  (from point  206  to point  207  in  FIG. 25 ). 
     The refrigerant flowing out of the mixing and pressurizing portion  16   e  of the ejector  16  flows into the first evaporator  17 . In the first evaporator  17 , low-pressure refrigerant is evaporated by absorbing heat from air blown by the blower fan  17   a,  so that the enthalpy of the refrigerant is increased (from point  207  to point  208  in  FIG. 25 ). Thus, air passing through the first evaporator  17  is cooled and the cooled air can be blown into the passenger compartment. The gas refrigerant flowing out of the first evaporator  17  is drawn into the first compressor  11  to be compressed again by the first compressor  11  (from point  208  to point  201  in  FIG. 25 ). 
     In contrast, the refrigerant stream flowing into the second passage  14   b  from the branch portion  13  is decompressed and expanded in iso-enthalpy by the throttle unit  18  (from point  203  to point  209  in  FIG. 25 ), and low-pressure refrigerant decompressed by the throttle unit  18  flows into the second evaporator  19 . In the second evaporator  19 , low-pressure refrigerant is evaporated by absorbing heat from air blown by the blower fan  19   a,  so that the enthalpy of the refrigerant is increased (from point  209  to point  210  in  FIG. 25 ). Thus, air passing through the second evaporator  19  is cooled so as to cool the interior of the cooler box. 
     In the fifteenth embodiment of the present invention, the throttled passage area of the throttle unit  18  can be set smaller than that of the first embodiment, thereby increasing the refrigerant decompression amount at the throttle unit  18 . Therefore, the refrigerant evaporation pressure (refrigerant evaporation temperature) in the second evaporator  19  can be set lower as compared with the first embodiment. 
     As shown in  FIG. 24 , the refrigerant flowing out of the second evaporator  19  is drawn into the second compressor  21 , and is compressed in the second compressor  21  (from point  210  to point  211  in  FIG. 25 ). Then, the compressed refrigerant is discharged from the second compressor  21  into the refrigerant suction port  16   d  of the ejector  16 , and is drawn into the mixing and pressurizing portion  16   e  of the ejector  16  from the refrigerant suction port  16   d.  Similarly to the first embodiment, the refrigerant is decompressed in iso-entropy while passing through the suction passage  16   i  (from point  211  to point  210 ′ in  FIG. 25 ). The other operations of the refrigeration cycle device  10  are similar to those of the above-described first embodiment. 
     In the refrigeration cycle device  10  having the ejector  16  according to the fifteenth embodiment, the refrigerant flowing out of the mixing and pressurizing portion  16   e  of the ejector  16  can be supplied to the first evaporator  17  while the refrigerant greatly decompressed by the throttle unit  18  in the second passage  14   b  can be supplied to the second evaporator  19  through the throttle unit  18 . Thus, both the first evaporator  17  and the second evaporator  19  can be operated simultaneously to have greatly different cooling capacities, and thereby the second evaporator  19  can be used to cool the interior of the cooler box that needs a cooling temperature lower than that in the passenger compartment. 
     At a low outside air temperature, a pressure difference between the refrigerant pressure on the high-pressure side and the refrigerant pressure on the low-pressure side becomes smaller in the refrigerant cycle of the refrigerant cycle device  10 . In this case, the flow amount of the refrigerant passing through the nozzle  16   a  of the ejector  16  may be decreased, and thereby the suction capacity of the ejector  13  may be decreased. Even in this case, because the second compressor  21  is located in the refrigeration cycle device  10  of the fifteenth embodiment, the suction capacity of the refrigerant into the ejector  16  from the refrigerant suction port  16   d  can be increased, so that the refrigerant cycle can be stably operated. 
     Furthermore, because the refrigerant is pressurized by using both the first and second compressors  11 ,  21 , a pressure difference between the suction pressure and the discharge pressure in respective compressors  11 ,  21  can be reduced. Thus, compression efficiency of each of the first and second compressors  11 ,  21  can be improved, thereby improving the COP in the refrigerant cycle of the refrigeration cycle device  10 . 
     The compression efficiency in the compressor  11 ,  21  is a ratio ΔE 1 /ΔE 2  of an increase amount ΔE 1  of the enthalpy of the refrigerant while being compressed in iso-entropy in the compressor  11 ,  21  to an increase amount ΔE 2  of the enthalpy of the refrigerant while being actually compressed in the compressor  11 ,  21 . For example, when the rotational speed or the pressurizing amount of the compressor  11 ,  21  is increased, the refrigerant temperature is increased by the friction force, and thereby the increase amount ΔE 2  is increased and the compression efficiency is reduced. 
     Thus, in the refrigeration cycle device  10 , if the pressure difference between the refrigerant pressure on the high-pressure side and the refrigerant pressure on the low-pressure side needs to be increased, the improving effect of the COP in the refrigerant cycle can be made significantly. 
     According to the fifteenth embodiment of the present invention, even when the ejector  16  is used for the refrigerant cycle device  10  provided with the first compressor  11  and the second compressor  21 , the ejector efficiency can be sufficiently improved. Furthermore, the refrigerant suction capacity of the ejector  16  can be suitably increased by using the second compressor  21 , and thereby the configuration of the ejector  16  can be easily set. 
     Thus, in the present embodiment, the ejector  16  can be easily configured so as to prevent the flow velocity of the suction refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  from being unnecessarily increased. That is, in the present embodiment, because the flow velocity of the refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  can be changed by not only the decompression characteristics in the suction passage  16   i  but also the discharge refrigerant pressure of the second compressor  21 , the suction passage  16   i  of the ejector  16  can be easily formed. Therefore, the flow velocity of the suction refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  can be easily adjusted by adjusting the refrigerant pressure at the refrigerant suction port  16   d  of the ejector  16 . 
     According to the fifteenth embodiment, by adjusting the refrigerant discharge capacity of the second compressor  21 , the flow velocity of the suction refrigerant flowing from the suction passage  16   i  into the mixing and pressurizing portion  16   e  can be easily adjusted at a suitable velocity, relative to the flow velocity of the jet refrigerant jetted from the refrigerant jet port  16   c  of the nozzle  16 . As a result, the configurations of respective parts in the ejector  16  can be easily set, and thereby the ejector  16  can be easily formed. 
     The Other Embodiments 
     Although the present invention has been fully described in connection with the preferred embodiments thereof with reference to the accompanying drawings, it is to be noted that various changes and modifications will become apparent to those skilled in the art. 
     According to any one embodiment of the present invention, the ejector  16  is provided with the suction passage  16   i  through which the refrigerant (fluid) drawn from the refrigerant suction port  16   d  flows into the inlet of the mixing and pressurizing portion  16   e.  The passage sectional area of the suction passage  16   i  is configured to be changed such that the refrigerant (fluid) drawn from the refrigerant suction port  16   d  is decompressed in the suction passage  16   i  substantially in iso-entropy. Alternatively, the passage area of the suction passage  16   i  is configured to be changed such that the flow velocity of the refrigerant flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is substantially equal to the flow velocity of the refrigerant (fluid) flowing from the jet port  16   c  of the nozzle  16   a  into the mixing and pressurizing portion  16   e.  Alternatively, the passage sectional area of the suction passage  16   i  is configured to be changed such that the flow velocity of the fluid flowing into the mixing and pressurizing portion  16   e  from the suction passage  16   i  is equal to or larger than the sound velocity. In this case, the ejector efficiency can be effectively improved. The other configurations in the ejector  16  may be suitably changed or combined without being limited to the above-described embodiments. 
     According to any one embodiment of the present invention, the mixing and pressurizing portion  16   e  is configured by the straight portion  16   g  extending from the inlet of the mixing and pressurizing portion  16   e  in a range in the axial direction, and the expanding portion  16   h  extending continuously from a downstream end of the straight portion  16   g  to the outlet of the mixing and pressurizing portion  16   e.  The straight portion  16   g  is a cylindrical passage having a constant passage area in its entire range, and the expending portion  16   h  is configured such that a passage sectional area of the expanding portion  16   h  is gradually increased toward downstream in the flow direction of the refrigerant. In the ejector  16 , the other configurations may be suitably changed or combined without being limited to the above-described embodiments. For example, the range of the straight portion  16   g  in the axial direction of the nozzle  16   a  is set such that the flow velocities of gas refrigerant and liquid refrigerant within the refrigerant flowing into the mixing and pressurizing portion  16   e  become equal to each other in the range. Alternatively, when the length of the straight portion  16   g  in the axial direction of the nozzle is L 1  and the length from the inlet of the mixing and pressurizing portion  16   e  to the outlet of the mixing and pressurizing portion  16   e  in the axial direction is L 2 , the mixing and pressurizing portion  16   e  is configured such that 0&lt;L 1 /L 2 ≦0.4. Alternatively, the mixing and pressurizing portion  16   e  may be configured such that the refrigerant is pressurized in iso-entropy therein. 
     In the above-described embodiments of the present invention, the ejector  16  is used for the refrigeration cycle device  10  in which the refrigerant is branched at the branch portion  13  on an upstream side of the nozzle  16   a  in the refrigerant flow from the radiator  12 . However, the ejector  16  of the present invention can be used for a refrigeration cycle device without being limited to the examples of the above-described embodiments. 
     For example, the ejector  16  of the present invention can be used for a refrigeration cycle device shown in  FIG. 26 . The refrigeration cycle device shown in  FIG. 26 , an accumulator  20  is located downstream of the outlet of the ejector  16  so that the refrigerant flowing out of the ejector  16  can directly flow into the accumulator  20 . The accumulator  20  has a gas refrigerant outlet coupled to the refrigerant suction side of the compressor  11 , and a liquid refrigerant outlet connected to a refrigerant inlet of an evaporator  19  so that the liquid refrigerant separated from the gas refrigerant in the accumulator  20  flows into the evaporator  19 . The gas refrigerant evaporated in the evaporator  19  is drawn into the refrigerant suction port  16   d  of the ejector  16 . In the refrigeration cycle device shown in  FIG. 26 , the refrigerant flowing from the radiator  12  is decompressed in the nozzle  16   a  and the gas refrigerant from the evaporator  19  is drawn into the ejector  16  from the refrigerant suction port  16   d  by the high-speed jet flow from the nozzle  16   a.  Even when the ejector  16  according to any one of the first to tenth embodiments is used for the refrigeration cycle device shown in  FIG. 26 , the ejector efficiency can be improved. 
     In the example shown in  FIGS. 5A and 5B , the suction gas refrigerant is decompressed in iso-entropy in the suction passage  16   i;  however, the suction gas refrigerant is not limited to be decompressed in iso-entropy. 
       FIGS. 27A and 27B  are modified examples of  FIG. 5B . X and Y in  FIG. 27A  and  FIG. 27B  correspond to the enlarged part VB in  FIG. 5A . As shown in  FIG. 27A , gas-liquid two-phase refrigerant can be drawn from the refrigerant suction port  16   b  and can be decompressed in iso-entropy in the suction passage  16   i  of the ejector  16 . Alternatively, as shown in  FIG. 27B , the gas refrigerant is drawn from the refrigerant suction port  16   d,  and can be decompressed in iso-entropy into a gas-liquid two-phase sate. 
     In the above-described embodiments, the Freon-based refrigerant or the carbon dioxide is typically used as the refrigerant. However, as the refrigerant, a generally-known refrigerant or a generally known fluid may be used. For example, carbon-hydride based refrigerant may be used as the refrigerant. 
     In the above-described embodiments, the refrigerant cycle device is used for a vehicle air conditioner or for a vehicle refrigerator. However, the refrigeration cycle device may be used for a fixed cooler, a fixed refrigerator, a box having a cooling function, a cooling device for a coin machine or the like. 
     In the above-described embodiments, the first and second evaporators  17 ,  19  are used as an interior heat exchanger for cooling air, and the radiator  12  is used as an exterior heat exchanger for radiating heat to outside air. However, the first and second evaporators  17 ,  19  may be used as an exterior heat exchanger for absorbing heat from outside air, and the radiator  12  may be used as an interior heat exchanger for heating a fluid to be heated such as water or air. That is, the ejector  16  of the present invention can be used for a heat pump cycle system with a heating function or/and a cooling function. 
     Such changes and modifications are to be understood as being within the scope of the present invention as defined by the appended claims.