US20090229306A1 - Vapor compression refrigerating cycle apparatus - Google Patents

Vapor compression refrigerating cycle apparatus Download PDF

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Publication number
US20090229306A1
US20090229306A1 US12/380,792 US38079209A US2009229306A1 US 20090229306 A1 US20090229306 A1 US 20090229306A1 US 38079209 A US38079209 A US 38079209A US 2009229306 A1 US2009229306 A1 US 2009229306A1
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United States
Prior art keywords
refrigerant
flow
nozzle portion
cycle apparatus
pressure
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Abandoned
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US12/380,792
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English (en)
Inventor
Etsuhisa Yamada
Haruyuki Nishijima
Gouta Ogata
Mika Gocho
Hideya Matsui
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Denso Corp
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Denso Corp
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Assigned to DENSO CORPORATION reassignment DENSO CORPORATION ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: GOCHO, MIKA, MATSUI, HIDEYA, NISHIJIMA, HARUYUKI, OGATA, GOUTA, YAMADA, ETSUHISA
Publication of US20090229306A1 publication Critical patent/US20090229306A1/en
Abandoned legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
    • F25B2341/001—Ejectors not being used as compression device
    • F25B2341/0011—Ejectors with the cooled primary flow at reduced or low pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2500/00—Problems to be solved
    • F25B2500/19—Calculation of parameters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B40/00—Subcoolers, desuperheaters or superheaters

Definitions

  • the present invention relates to a vapor compression refrigerating cycle apparatus having an ejector as a refrigerant decompressing and circulating device.
  • a vapor compression refrigerating cycle apparatus is, for example, described in JP-A-2007-23966 (US2006/0266072 A1).
  • the described refrigerating cycle apparatus has an ejector as a decompressing device for decompressing condensed refrigerant and two evaporators.
  • the ejector generally has a nozzle portion, a suction portion, a mixing portion and a pressure-increase portion.
  • the nozzle portion draws a part of the refrigerant downstream of a radiator, and decompresses and expands the drawn refrigerant in an isenthalpic manner.
  • the suction portion draws a remaining part of the refrigerant from one of the evaporators.
  • the part of the refrigerant is jetted from the nozzle portion at high velocity, and is mixed with the remaining part of the refrigerant drawn from the suction portion. Further, the mixed refrigerant is increased in pressure through the pressure-increase portion, and is then discharged from the ejector.
  • the refrigerant is further conducted to the other evaporator to be evaporated, and is then drawn into the compressor.
  • the present invention is made in view of the foregoing matter, and it is an object of the present invention to provide a vapor compression refrigerating cycle apparatus capable of controlling a condition of refrigerant at the nozzle portion of the ejector to a predetermined condition, thereby to maintain the COP at the sufficient level.
  • a vapor compression refrigerating cycle apparatus includes a compressor, a radiator, first and second decompressing devices, a flow distributor, an ejector and a suction-side evaporator.
  • the compressor draws and compresses refrigerant.
  • the radiator radiates heat of refrigerant discharged from the compressor.
  • the first decompressing device decompresses refrigerant discharged from the radiator.
  • the flow distributor separates refrigerant decompressed by the first decompressing device into at least a first flow and a second flow.
  • the ejector includes a nozzle portion and a suction portion.
  • the nozzle portion draws refrigerant of the first flow, and decompresses and expands the refrigerant of the first flow to generate a refrigerant jet flow.
  • the suction portion draws refrigerant of the second flow by the refrigerant jet flow from the nozzle portion.
  • the second decompressing device decompresses the refrigerant of the second flow.
  • the suction-side evaporator evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector.
  • the vapor compression refrigerating cycle apparatus is configured such that refrigerant pressure (P 0 ) at an inlet of the first decompressing device, refrigerant pressure (P) at an inlet of the nozzle portion, refrigerant pressure (P 2 ) at an outlet of the nozzle portion satisfy a pressure relationship of 0.1 ⁇ (P 0 ⁇ P 2 )>(P 0 ⁇ P) ⁇ 0.6 ⁇ (P 0 ⁇ P 2 ):
  • the refrigerant pressure at the inlet of the nozzle portion becomes an optimum condition
  • a distribution ratio of the refrigerant to the suction-side evaporator and the nozzle portion can be set to an optimum ratio. Therefore, capacity of the suction-side evaporator and nozzle efficiency are both improved.
  • the COP of the vapor compression refrigerating cycle apparatus improves.
  • the pressure relationship is achieved by adjusting a throttle degree of at least one of the first decompressing device, the second decompressing device and the nozzle portion.
  • a vapor compression refrigerating cycle apparatus includes a compressor, a radiator, first and second decompressing devices, a flow distributor, an ejector and a suction-side evaporator.
  • the compressor draws and compresses refrigerant.
  • the radiator radiates heat of refrigerant discharged from the compressor.
  • the first decompressing device decompresses refrigerant discharged from the radiator.
  • the flow distributor separates refrigerant decompressed by the first decompressing device into at least a first flow and a second flow.
  • the ejector includes a nozzle portion and a suction portion.
  • the nozzle portion draws refrigerant of the first flow, and decompresses and expands the refrigerant of the first flow to generate a refrigerant jet flow.
  • the suction portion draws refrigerant of the second flow by the refrigerant jet flow from the nozzle portion.
  • the second decompressing device decompresses the refrigerant of the second flow.
  • the suction-side evaporator evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector.
  • the vapor compression refrigerating cycle apparatus is configured such that the refrigerant at an inlet of the nozzle portion has a dryness in a range between 0.003 and 0.14.
  • a distribution ratio of the refrigerant to the suction-side evaporator and the nozzle portion can be set to an optimum ratio. Therefore, capacity of the suction-side evaporator and nozzle efficiency are both improved. As such, the COP of the vapor compression refrigerating cycle apparatus improves.
  • FIG. 1 is a schematic block diagram of a vapor compression refrigerating cycle apparatus according to an embodiment of the present invention
  • FIG. 2 is a graph showing a relationship between enthalpy and pressure in the vapor compression refrigerating cycle apparatus according to the embodiment
  • FIG. 3 is a graph showing an operation of the vapor compression refrigerating cycle apparatus according to the embodiment.
  • FIG. 4 is a graph showing a relationship between refrigerant pressure and a COP improvement effect of the vapor compression refrigerating cycle apparatus according to the embodiment
  • FIG. 5 is a graph showing a relationship between dryness of refrigerant at an inlet of a nozzle portion of an ejector and the COP improvement effect of the vapor compression refrigerating cycle apparatus according to the embodiment;
  • FIG. 6 is a schematic block diagram of a vapor compression refrigerating cycle apparatus according to another embodiment of the present invention.
  • FIG. 7 is a graph showing a relationship between enthalpy and pressure in the vapor compression refrigerating cycle apparatus shown in FIG. 6 .
  • FIG. 1 shows an example of a vapor compression refrigerating cycle apparatus 10 of the present embodiment.
  • the refrigerating cycle apparatus 10 is an ejector-type refrigerating cycle apparatus including an ejector 5 , which serves as a decompressing device for decompressing refrigerant and a pump for transporting the refrigerant.
  • the refrigerating cycle apparatus 10 is, for example, employed in a vehicle refrigerating unit, a vehicle air conditioner and the like.
  • the refrigerating cycle apparatus 10 generally includes a compressor 1 , a radiator 2 , a first decompressing device 3 , an ejector 5 , a flow distributor 6 , a second decompressing device 4 and a suction-side evaporator 8 .
  • the refrigerating cycle apparatus 10 further includes a discharge-side evaporator 7 .
  • the compressor 1 , the radiator 2 , the first decompressing device 3 , the ejector 5 , the flow distributor 6 and the discharge-side evaporator 7 (hereinafter, referred to as the first evaporator 7 ) are connected in a form of loop through pipes.
  • the flow distributor 6 distributes the refrigerant, which has been decompressed through the first decompressing device 3 into a first flow that is in communication with a nozzle portion 5 a of the ejector 5 and a second flow that is in communication with a suction portion 5 b of the ejector 5 through a branch passage 9 . That is, the branch passage 9 diverges from the flow distributor 6 and connects to the suction portion 5 b of the ejector 5 .
  • the second decompressing device 4 and the suction-side evaporator 8 (hereinafter, referred to as the second evaporator 8 ) are disposed on the branch passage 9 .
  • the compressor 1 draws and compresses refrigerant.
  • the compressor 1 discharges high pressure refrigerant toward the radiator 2 .
  • the compressor 1 is driven by a vehicle engine through an electromagnetic clutch, a pulley, and a belt.
  • the compressor 1 is any-types of compressor, such as, a variable capacity-type compressor that is capable of adjusting a discharge rate in accordance with a change in discharge capacity, a fixed capacity-type compressor that is capable of adjusting a discharge rate in accordance with a change in a rate of operation thereof by on and off operations of the electromagnetic clutch, an electric compressor that is capable of adjusting a discharge rate by controlling a rotation speed of an electric motor, or the like.
  • the radiator 2 is disposed downstream of the compressor 1 with respect to a flow of refrigerant.
  • the radiator 2 performs heat exchange between the high pressure refrigerant discharged from the compressor 1 and air, thereby to condense the refrigerant.
  • the air is, for example, outside air drawn from an outside of a passenger compartment of a vehicle and forcibly applied to the radiator 2 such as by a blower (not shown).
  • the refrigerant is not limited to a specific refrigerant.
  • the refrigerant is R404A.
  • the refrigerating cycle apparatus is operated under a subcritical condition where pressure on a high-pressure side does not exceed the critical pressure.
  • the radiator 2 serves as a condenser for condensing the refrigerant therein.
  • the refrigerating cycle apparatus is operated under a supercritical condition where pressure on the high-pressure side exceeds the critical pressure. In this case, the refrigerant radiates heat while maintaining in a supercritical condition, and thus is not condensed.
  • the first decompressing device 3 serves to decompress the high pressure refrigerant having passed through the radiator 2 .
  • the first decompressing device 3 is, for example, an expansion valve.
  • the expansion valve 3 is, for example, a temperature operation-type in which a valve opening degree is controlled to adjust a superheat degree to a predetermined condition based on a temperature of refrigerant at an outlet of the first evaporator 7 .
  • the first decompressing device 3 can be a fixed flow control valve, an electric controlled flow control valve in which a refrigerant flow rate is variably controlled, or the like.
  • the high pressure refrigerant is decompressed into a gas and liquid two-phase condition by controlling a decompressing rate through the first decompressing device 3 , and is then conducted to the flow distributor 6 .
  • the gas and liquid two-phase refrigerant forms stratified flow, linear flow, slag flow, and the like in accordance with dryness, velocity and the like. Further, the gas and liquid two-phase refrigerant forms an upper and lower separated flow in which gas-phase refrigerant is located above liquid refrigerant.
  • the flow distributor 6 is a block member having such as a cubic shape and a rectangular parallelepiped shape.
  • the flow distributor 6 is formed with multiple passages therein, and serves to distribute the refrigerant decompressed through the first decompressing device 3 into at least two flows at predetermined rates.
  • the flow distributor 6 at least has a first passage that is in communication with the first decompressing device 3 , a second passage diverging from the first passage and connecting to the branch passage 9 for conducting the refrigerant toward the second evaporator 8 , and a third passage diverging from the first passage and is in communication with the nozzle portion 5 a of the ejector 5 .
  • the first to third passages constitute a distribution rate adjusting part.
  • Each of the first to third passages has a predetermined shape and a passage area (cross-sectional area) and is located at a predetermined position, such as a predetermined height.
  • the passage areas of the first to third passages satisfy a predetermined relationship. Therefore, the flow rate of refrigerant passing through each passage, the volume of liquid-phase refrigerant passing through each passage and the like are determined in accordance with pressure condition of the refrigerant.
  • the low distributor 6 can be provided with a valve device to vary the flow rates of the refrigerant passing through the respective passages.
  • the ejector 5 serves as a decompressing device for decompressing refrigerant and a circulating device for circulating the refrigerant by means of a drawing effect (dragging effect) generated by a jet flow of refrigerant.
  • the ejector 5 generally has the nozzle portion 5 a, the suction portion 5 b, a mixing portion 5 c and a diffuser portion 5 d.
  • the nozzle portion 5 a is in communication with the third passage of the flow distributor 6 .
  • the nozzle portion 5 a draws the refrigerant of the first flow from the flow distributor 6 , and decompresses and expands the refrigerant in an isenthalpic manner by reducing a passage area therein.
  • the suction portion 5 b is disposed to be in communication with a jet port of the nozzle portion 5 a.
  • the suction portion 5 b draws gas-phase refrigerant from the second evaporator 8 .
  • the mixing portion 5 c mixes the refrigerant jetted from the jet port of the nozzle portion 5 a at high velocity with the refrigerant drawn from the suction portion 5 b.
  • the diffuser portion 5 d is disposed downstream of the mixing portion 5 c.
  • the diffuser portion 5 d is configured such that a passage area gradually reduces to reduce the velocity of the refrigerant and increase the refrigerant in pressure. That is, the diffuser portion 5 d has a function of converting velocity energy of the refrigerant into pressure energy. Therefore, the diffuser portion 5 d can be also referred to as a pressure-increase portion.
  • pressure is rapidly reduced in the nozzle portion 5 a, and is the lowest at the outlet of the nozzle portion 5 a. Since the refrigerant decompressed in the nozzle portion 5 a is mixed with the refrigerant drawn from the suction portion 5 b in the mixing portion 5 c, the pressure gradually increases. The pressure is then increased in the diffuser portion 5 d due to the decrease in velocity.
  • the first evaporator 7 is disposed downstream of the diffuser portion 5 d with respect to the flow of refrigerant.
  • the first evaporator 7 is a heat absorber that performs heat exchange between the refrigerant discharged from the ejector 5 and air, which is forcibly applied to the first evaporator 7 , thereby to achieve a heat absorbing effect due to evaporation of the refrigerant.
  • a discharge side of the first evaporator 7 is in communication with a suction side of the compressor 1 .
  • the second decompressing device 4 is, for example, constructed of a capillary tube, such as a spiral tubule.
  • the second decompressing device 4 is disposed on the branch passage 9 .
  • the second decompressing device 4 serves to decompress refrigerant flowing in the second evaporator 8 and control a flow rate of the refrigerant.
  • the second decompressing device 4 can be a variable decompressing device such as an electric control expansion valve, in place of the capillary tube.
  • the second evaporator 8 is disposed on the branch passage 9 downstream of the second decompressing device 4 with respect to the flow of refrigerant.
  • the second evaporator 8 is a heat absorber, similar to the first evaporator 7 . That is, the second evaporator 8 achieves a heat absorbing effect by evaporating the refrigerant.
  • the second evaporator 8 is located downstream of the first evaporator 7 with respect to the flow of air.
  • the air having passed through the first evaporator 7 is further cooled while passing through the second evaporator 8 by exchanging heat with the refrigerant flowing inside of the second evaporator 8 .
  • the air is conducted to a predetermined space, such as for an air conditioning operation.
  • first evaporator 7 and the second evaporator 8 are provided differently.
  • airs can be applied separately to the first evaporator 7 and the second evaporator 8 by blowers and the like, and the airs can be conducted to different spaces to be air-conditioned.
  • the first evaporator 7 and the second evaporator 8 can be constructed separately from each other. Alternatively, the first evaporator 7 and the second evaporator 8 can be integrated with each other. In a case where the first evaporator 7 and the second evaporator 8 are integrated with each other, the first evaporator 7 and the second evaporator 8 can be joined with each other by brazing. In this case, components of the first evaporator 7 and the second evaporator 8 are made of aluminum, for example. Further, the flow distributor 6 , the second decompressing device 4 and the ejector 5 can be integrated with each other into a unit, and further fixed to the first and second evaporators 7 , 8 .
  • the vapor compression refrigerating cycle apparatus 10 can be further provided with an internal heat exchanger to perform heat exchange between the high pressure refrigerant flowing between the radiator 2 and the first decompressing device 3 and low pressure refrigerant to be drawn to the compressor 1 .
  • the high pressure refrigerant flowing between the radiator 2 and the expansion valve 3 is cooled by the heat exchange with the low pressure refrigerant.
  • enthalpy differential between refrigerant inlets and refrigerant outlets of the first evaporator 7 and the second evaporator 8 increases, and thus cooling capacity improves.
  • the control unit is constructed of a microcomputer including a CPU, a ROM, a RAM and the like and peripheral circuits.
  • the control unit executes various computations and processing in accordance with control programs stored in the ROM to control operations of various devices including the compressor 1 .
  • the control unit receives detection signals from various sensors and various manipulation signals from an operation panel (not shown).
  • the operation panel is provided with a temperature setting switch for setting a cooling temperature of a space to be cooled and an air conditioner operation switch for generating an operation command signal of the compressor 1 .
  • points a 1 through i 1 correspond to locations a 1 through i 1 in FIG. 1 .
  • the electromagnetic clutch of the compressor 1 When the electromagnetic clutch of the compressor 1 is electrically conducted in accordance with the signal generated from the control unit, the electromagnetic clutch becomes in a connected state and a driving force is transmitted from an engine of a vehicle to the compressor 1 .
  • the operation of the compressor 1 When the operation of the compressor 1 is started, the gas-phase refrigerant is drawn into the compressor 1 from the first evaporator 7 and compressed in the compressor 1 .
  • the high temperature, high pressure refrigerant is condensed by being cooled by the air. (a 1 ⁇ b 1 )
  • High pressure liquid-phase refrigerant flowing out from the radiator 2 at the flow rate G is decompressed and expanded into predetermined pressure by the first decompressing device 3 .
  • the gas and liquid two-phase refrigerant is generated.
  • refrigerant pressure at an inlet of the first decompressing device 3 is defined as P 0 .
  • the gas and liquid two-phase refrigerant flowing out from the first decompressing device 3 flows in the flow distributor 6 .
  • the gas and liquid two-phase refrigerant is separated into the first flow passing through the third passage toward the nozzle portion 5 a of the ejector 5 (b 1 ⁇ c 1 ) and the second flow passing through the second passage toward the second decompressing device 4 (b 1 ⁇ h 1 ), at predetermined flow rates.
  • the flow rate of the refrigerant of the first flow is defined as Gn
  • the flow rate of the refrigerant of the second flow is defined as Ge.
  • Refrigerant pressure at an inlet of the nozzle portion 5 a is defined as P.
  • the refrigerant flows in the nozzle portion 5 a of the ejector 5 at the flow rate Gn from the first flow.
  • the refrigerant is decompressed and expanded in the isenthalpic manner through the nozzle portion 5 a. (c 1 ⁇ d 1 ).
  • the refrigerant pressure P reduces to refrigerant pressure P 2 at the outlet of the nozzle portion 5 a. That is, in the nozzle portion 5 a, pressure energy of the refrigerant is converted into velocity energy, and thus the refrigerant is jetted from the jet port of the nozzle portion 5 a at high velocity.
  • the gas-phase refrigerant of the flow rate Ge is drawn from the second evaporator 8 into the suction portion 5 b by the drawing effect generated by the jet flow of the refrigerant.
  • the refrigerant jetted from the nozzle portion 5 a and the refrigerant drawn into the suction portion 5 b are mixed with each other in the mixing portion 5 c (d 1 ⁇ e 1 , i 1 ⁇ e 1 ), and then introduced in the diffuser portion 5 d.
  • the diffuser portion 5 d since the passage area is gradually increased, velocity (expansion) energy of the refrigerant is converted into pressure energy. Thus, the refrigerant is increased in pressure (e 1 ⁇ f 1 ).
  • the low temperature, low pressure refrigerant is evaporated in a heat exchanging core portion by absorbing heat from the air (f 1 ⁇ g 1 ). Pressure of the low temperature, low pressure refrigerant is defined as P 1 .
  • the gas-phase refrigerant evaporated in the first evaporator 1 is drawn by the compressor 1 and is compressed again.
  • the refrigerant of the second flow is conducted in the branch passage 9 at the flow rate Ge and decompressed into the low pressure refrigerant by the second decompressing device 4 (b 1 ⁇ h 1 ).
  • the low pressure refrigerant is then conducted to the second evaporator 8 .
  • the low pressure refrigerant is evaporated by absorbing heat from the air (h 1 ⁇ i 1 ), and becomes the gas-phase refrigerant.
  • the gas-phase refrigerant is drawn into the suction portion 5 b at the flow rate Ge.
  • the refrigerant of the flow rate Gn is supplied to the first evaporator 7 and the refrigerant of the flow rate Ge is supplied to the second evaporator 8 through the second decompressing device 4 . Therefore, cooling effects are achieved simultaneously by the first and second evaporators 7 , 8 .
  • the first decompressing device 3 , the second decompressing device 4 and the nozzle portion 5 a have predetermined throttle degrees such that the refrigerant pressure P 0 at the inlet of the first decompressing device 3 , the refrigerant pressure P at the inlet of the nozzle portion 5 a and the refrigerant pressure P 2 at the outlet of the nozzle portion 5 a satisfy the following pressure relationship (R1):
  • the vapor compression refrigerating cycle apparatus 10 is configured such that a decrease in pressure, that is, a differential pressure between the refrigerant pressure P 0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to a value that is obtained by multiplying a differential pressure between the inlet of the first decompressing device 3 and the outlet of the nozzle portion 5 a by a value that is at least 0.1 and at most 0.6.
  • ⁇ P represents an increase in pressure by the ejector 5 , such as by the diffuser portion 5 d. That is, ⁇ P is a differential pressure (P 1 ⁇ P 2 ) between the refrigerant pressure P 1 flowing in the first evaporator 7 and a refrigerant evaporation pressure P 2 in the second evaporator 8 . Because suction pressure of the compressor 1 is increased by an effect of increasing in pressure by the diffuser portion 5 d, which is represented by ⁇ P, the driving force of the compressor 1 can be reduced. As a result, the COP of the vapor compression refrigerating cycle apparatus 10 improves.
  • the refrigerant evaporation pressure P 2 of the second evaporator 8 is lower than the refrigerant evaporation pressure P 1 of the first evaporator 7 . Therefore, a refrigerant evaporation temperature of the second evaporator 8 is lower than a refrigerant evaporation temperature of the first evaporator 7 .
  • first evaporator 7 is disposed upstream of the second evaporator 8 with respect to the flow of air, it is possible to ensure both a temperature differential between the refrigerant evaporation temperature of the first evaporator 7 and the air and a temperature differential between the refrigerant evaporation temperature of the second evaporator 8 and the air. Accordingly, cooling performances of both the first and second evaporators 7 , 8 effectively improve.
  • FIG. 3 shows relationships between differential pressure at inlets and outlets of flow rate control devices, such as the first decompressing device 3 , the second decompressing device 4 and the nozzle portion 5 a, and the flow rates at the respective portions.
  • the flow rate G of the first decompressing device 3 increases as the refrigerant pressure P at the inlet of the nozzle portion 5 a reduces, that is, as the differential pressure (P 0 ⁇ P) between the refrigerant pressure P 0 of the inlet of the first decompressing device 3 and the refrigerant pressure P of the inlet of the nozzle portion 5 a increases.
  • the differential pressure between the inlet and the outlet of each of the nozzle portion 5 a and the second decompressing device 4 reduces.
  • each of the flow rates Gn, Ge reduces.
  • the refrigerant pressure P at the inlet of the nozzle portion 5 a is determined to pressure where the flow rate G of the first decompressing device 3 is equal to the sum of the flow rate Gn of the nozzle portion 5 a and the flow rate Ge of the second decompressing device 4 .
  • a ratio of the flow rates Gn, Ge is determined based on a flow rate property by the differential pressure between the inlet and the outlet of the nozzle portion 5 a and a flow rate property by the differential pressure between the inlet and the outlet of the second decompressing device 4 . Further, expansion energy recovered at the nozzle portion 5 a reduces as the refrigerant pressure P at the inlet of the nozzle portion 5 a reduces. As such, the increase in pressure ⁇ P by the ejector 5 reduces.
  • the ratio of the flow rates Gn, Ge it is preferable to set the ratio of the flow rates Gn, Ge to an optimum ratio as discussed hereinabove, and it is recognized that there is an optimum condition of the refrigerant pressure at the inlet of the nozzle portion 5 a. Further, it is realized that the nozzle efficiency is sufficient when the pressure relationship (R1) is satisfied because the pressure condition at the inlet of the nozzle portion 5 a is under the optimum condition. Moreover, it is realized that the refrigerating capacity (COP) is sufficiently achieved in a range of the refrigerant flow rate ratio, which is obtained when the pressure relationship (R1) is satisfied. The range of the refrigerant flow rate ratio corresponds to a nondimensional flow rate ratio (Ge/(Ge+Gn)).
  • FIG. 4 shows a relationship between a pressure ratio (P 0 ⁇ P)/(P 0 ⁇ P 2 ) and a COP improvement effect.
  • the pressure ratio (P 0 ⁇ P)/(P 0 ⁇ P 2 ) is a ratio of the decrease in the refrigerant pressure P at the inlet of the nozzle portion 5 a with respect to the refrigerant pressure P 0 at the inlet of the first decompressing device 3 to the decrease in refrigerant pressure P 2 at the outlet of the nozzle portion 5 a with respect to the refrigerant pressure P 0 at the inlet of the first decompressing device 3 .
  • the COP improvement effect is the improvement of the COP of the vapor compression refrigerating cycle apparatus 10 with respect to the COP of an expansion valve cycle apparatus. That is, the higher the value indicative of the COP improvement effect is, the more the COP of the vapor compression refrigerating cycle apparatus 10 is improved, as compared with the COP of the expansion valve cycle apparatus.
  • the expansion valve cycle apparatus is a refrigerating cycle apparatus constructed by sequentially connecting the compressor, the radiator, the expansion valve and the evaporator in a form of closed circuit.
  • the COP improvement effect is low in regions where the pressure ratio (P 0 ⁇ P)/(P 0 ⁇ P 2 ) is small and large. Further, the COP improvement effect is high in a middle region between the regions. Particularly, when the pressure ratio (P 0 ⁇ P)/(P 0 ⁇ P 2 ) is in a range between 0.1 and 0.6, the COP improvement effect is stable and at the highest level. That is, the pressure ratio (P 0 ⁇ P)/(P 0 ⁇ P 2 ) is the optimum in the range between 0. and 0.6.
  • the COP which is obtained by a ratio of the refrigerating capacity Qer of the entirety of the refrigerating cycle apparatus to the driving force L of the compressor 1 , reduces.
  • the pressure relationship (R1) is achieved by constructing the first decompressing device 3 , the second decompressing device 4 and the ejector 5 to have the predetermined throttle degrees, respectively.
  • the refrigerant at the inlet of the nozzle portion 5 a is controlled under a predetermined pressure condition. Accordingly, the COP is sufficiently ensured.
  • FIG. 5 is a graph showing a relationship between dryness X of the refrigerant at the inlet of the nozzle portion 5 a and the COP improvement effect of the vapor compression refrigerating cycle apparatus 10 .
  • the dryness X is a ratio of vapor in 1 kg wet vapor of the refrigerant at the inlet of the nozzle portion 5 a. That is, the dryness X means that refrigerant contains X kg of dry saturated vapor and (1 ⁇ X) kg of saturated liquid.
  • the COP improvement effect means the improvement of the COP of the vapor compression refrigerating cycle apparatus 10 with respect to the COP of the expansion valve cycle apparatus, similar to FIG. 4 . That is, the higher the value of the COP improvement effect is, the more the COP of the vapor compression refrigerating cycle apparatus 10 is improved, as compared with the COP of the expansion valve cycle apparatus.
  • the COP improvement effect is low in regions where the dryness X is small and large.
  • the COP improvement effect is high in the middle region.
  • the COP improvement effect is stable and at the maximum level. That is, the dryness X is the optimum in the range between 0.003 and 0.14.
  • the nozzle efficiency is sufficiently ensured when the dryness X is in the range between 0.003 and 0.14, similar to FIG. 3 . In this case, however, the nozzle efficiency has a peak on a side adjacent to 0.003.
  • the refrigerant pressure at the inlet of the nozzle portion 5 a can be maintained to an optimum condition, similar to FIG. 3 , in accordance with the flow rate properties of the nozzle portion 5 a and the second decompressing device 4 . Therefore, the refrigerating capacity of the evaporators 7 , 8 and the increase in pressure ⁇ P by the ejector 5 are ensured in a balanced condition. As such, the COP of the refrigerating cycle apparatus 10 is sufficiently increased, as compared with the expansion valve cycle apparatus.
  • the dryness X of the refrigerant at the inlet of the nozzle portion 5 a can be controlled in the range between 0.003 and 0.14 by setting the throttle degrees of the second decompressing device 4 and the ejector 5 are to predetermined degrees. That is, by setting the throttle degrees of the second decompressing device 4 and the ejector 5 to the predetermined degrees, the refrigerant at the inlet of the nozzle portion 5 a can be controlled to a predetermined condition, such as equivalent to the condition shown in FIG. 3 . Therefore, the COP of the refrigerating cycle apparatus 10 improves.
  • the vapor compression refrigerating cycle apparatus 10 is configured such that the differential (P 0 ⁇ P) between the refrigerant pressure P 0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to the value that is obtained by multiplying the differential (P 0 ⁇ P 2 ) between the refrigerant pressure P 0 and the refrigerant pressure P 2 at the outlet of the nozzle portion 5 a by the value that is at least 0.1 and at most 0.6.
  • the above pressure relationship (R1) can be achieved by setting at least one of the throttle degrees of the first decompressing device 3 , the second decompressing device 4 and the nozzle portion 5 a to the predetermined degrees, for example.
  • the distribution ratio of the refrigerant to the second evaporator 8 and the nozzle portion 5 a becomes an optimum condition. Therefore, the performance of the evaporators 7 , 8 and the efficiency of the ejector 5 , such as the nozzle efficiency and the ejector efficiency, can be both ensured. Accordingly, the COP of the refrigerating cycle apparatus 10 improves, as compared with the expansion valve cycle apparatus.
  • the above example can be employed to the vapor compression refrigerating cycle apparatus including at least the compressor 1 , the radiator 2 , the first decompressing device 3 , the flow distributor 6 , the ejector 5 , the second decompressing device 4 , and the suction-side evaporator 8 . That is, even in the vapor compression refrigerating cycle apparatus without having the first evaporator 7 , it can be configured to have the pressure relationship (R1). Also in this vapor compression refrigerating cycle apparatus, the similar effects are achieved.
  • the vapor compression refrigerating cycle apparatus 10 is configured such that the dryness X of the refrigerant at the inlet of the nozzle portion 5 a is in the range between 0.003 and 0.14.
  • the dryness X in the range between 0.003 and 0.14 is achieved by setting at least one of the throttle degrees of the first decompressing device 3 , the second decompressing device 4 and the nozzle portion 5 a to the predetermined degree.
  • the COP of the refrigerating cycle apparatus 10 improves, as compared with the expansion valve cycle apparatus.
  • the above example can be employed to the vapor compression refrigerating cycle apparatus including at least the compressor 1 , the radiator 2 , the first decompressing device 3 , the flow distributor 6 , the ejector 5 , the second decompressing device 4 , and the suction-side evaporator 8 . That is, even in the vapor compression refrigerating cycle apparatus without having the first evaporator 7 , it can be configured such that the refrigerant has the dryness X in the above range at the inlet of the nozzle portion 5 a. Also in this vapor compression refrigerating cycle apparatus, the similar effects are achieved.
  • the vapor compression refrigerating cycle apparatus 10 can be configured such that the differential (P 0 ⁇ P) between the refrigerant pressure P 0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to the value that is obtained by multiplying the differential (P 0 ⁇ P 2 ) between the refrigerant pressure P 0 and the refrigerant pressure P 2 at the outlet of the nozzle portion 5 a by the value that is at least 0.1 and at most 0.6, and the dryness of the refrigerant at the inlet of the nozzle portion 5 a is in the range between 0.003 and 0.14.
  • the vapor compression refrigerating cycle apparatus 10 can be operated while appropriately maintaining the pressure and enthalpy. Accordingly, the performance of the evaporators 7 , 8 and the efficiency of the ejector 5 further improves, and the COP further improves.
  • the dryness X of the refrigerant at the inlet of the nozzle portion 5 a can be adjusted by the distribution rate adjusting means of the flow distributor 6 .
  • the mixing ratio of the liquid-phase refrigerant and the gas-phase refrigerant flowing toward the nozzle portion 5 a is controlled. Therefore, the dryness X can be more precisely adjusted.
  • the vapor compression refrigerating cycle apparatus 10 can be further modified as follows.
  • the discharge-side evaporator 7 can be eliminated and an internal heat exchanger 70 that performs heat exchange between the high pressure refrigerant discharged from the radiator 2 and the low pressure refrigerant discharged from the ejector 5 can be added.
  • the enthalpy of the low pressure refrigerant discharged from the ejector 5 can be increased from the point f 1 to the point g 1 , and the enthalpy of the refrigerant flowing into the suction-side evaporator 8 can be decreased from the point b 1 to a point b′ 1 .
  • the capacity of the suction-side evaporator 8 increases.
  • the refrigerating cycle apparatus can be configured to satisfy one of or both of the pressure relationship (R1) and the above optimum range of the dryness X.
  • the discharge-side evaporator 7 can be eliminated, and an accumulator as a low pressure-side gas and liquid separator for separating the refrigerant discharged from the ejector 5 into the gas-phase refrigerant and the liquid-phase refrigerant can be added.
  • the refrigerating cycle apparatus can be configured to satisfy one of or both of the pressure relationship (R1) and the above optimum range of the dryness X. Thus, the similar effects can be achieved.
  • the vapor compression refrigerating cycle apparatus 10 as discussed hereinabove can be employed to a heat pump cycle such as for a hot water supply apparatus or an interior air conditioner, and mounted in a movable unit such as a vehicle or in a fixed unit fixed at a predetermined location.
  • the refrigerant is not limited to R404 refrigerant.
  • the refrigerant can be any other types, such as chlorofluorocarbon-base refrigerant, HC-base refrigerant, carbon dioxide refrigerant or the like, which can be used in the supercritical cycle and the subcritical cycle. Even when the refrigerant other than R404 is used, the similar effects can be achieved.
  • the pressure relationship (R1) can be achieved by various ways, instead by setting the throttle degree of at least one of the first decompressing device 3 , the second decompressing device 4 , and the fixed nozzle portion 5 a to the predetermined degree.
  • the ejector 5 has a flow rate variable nozzle portion in which the throttle degree of the nozzle portion is variable in accordance with the movement of a valve rod, in place of the fixed nozzle portion 5 a.
  • the pressure relationship (R1) can be satisfied by adjusting the throttle degree of the nozzle portion.
  • the second decompressing device 4 can be constructed of a flow rate control variable-type decompressing device such as an electric control expansion valve, in place of the capillary tube 4 .
  • the pressure relationship (R1) can be achieved by adjusting the throttle degree of the second decompressing device 4 .
  • An operation of the flow rate control variable-type decompressing device is, for example, controlled by the control unit.
  • the flow distributor 6 is not limited to the block member having the passages therein, but can be constructed of any other types of distributors.
  • the flow distributor 6 can be constructed of a manifold pipe having branched passages.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
  • Jet Pumps And Other Pumps (AREA)
  • Air-Conditioning For Vehicles (AREA)
US12/380,792 2008-03-13 2009-03-04 Vapor compression refrigerating cycle apparatus Abandoned US20090229306A1 (en)

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JP2008064665A JP2009222255A (ja) 2008-03-13 2008-03-13 蒸気圧縮式冷凍サイクル

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US9261298B2 (en) 2010-07-23 2016-02-16 Carrier Corporation Ejector cycle refrigerant separator

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JP6116810B2 (ja) * 2012-03-23 2017-04-19 株式会社デンソー 冷凍サイクル装置
WO2013160929A1 (ja) * 2012-04-23 2013-10-31 三菱電機株式会社 冷凍サイクルシステム
CN102778076B (zh) * 2012-07-12 2014-07-30 西安交通大学 一种用于双温电冰箱的新型压缩/喷射混合制冷循环***
JP6102552B2 (ja) * 2012-11-16 2017-03-29 株式会社デンソー 冷凍サイクル装置

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CN101532740B (zh) 2011-05-04
DE102009012360A1 (de) 2009-10-01
JP2009222255A (ja) 2009-10-01

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