EP2995885B1 - Dispositif de réfrigération binaire - Google Patents

Dispositif de réfrigération binaire Download PDF

Info

Publication number
EP2995885B1
EP2995885B1 EP13883949.3A EP13883949A EP2995885B1 EP 2995885 B1 EP2995885 B1 EP 2995885B1 EP 13883949 A EP13883949 A EP 13883949A EP 2995885 B1 EP2995885 B1 EP 2995885B1
Authority
EP
European Patent Office
Prior art keywords
stage
refrigeration cycle
refrigerant
outdoor
stage refrigeration
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Active
Application number
EP13883949.3A
Other languages
German (de)
English (en)
Other versions
EP2995885A1 (fr
EP2995885A4 (fr
Inventor
Tomotaka Ishikawa
So Nomoto
Takeshi Sugimoto
Tetsuya Yamashita
Takashi Ikeda
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Mitsubishi Electric Corp
Original Assignee
Mitsubishi Electric Corp
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Mitsubishi Electric Corp filed Critical Mitsubishi Electric Corp
Publication of EP2995885A1 publication Critical patent/EP2995885A1/fr
Publication of EP2995885A4 publication Critical patent/EP2995885A4/fr
Application granted granted Critical
Publication of EP2995885B1 publication Critical patent/EP2995885B1/fr
Active legal-status Critical Current
Anticipated expiration legal-status Critical

Links

Images

Classifications

    • 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
    • F25B6/00—Compression machines, plants or systems, with several condenser circuits
    • F25B6/04—Compression machines, plants or systems, with several condenser circuits arranged in series
    • 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
    • F25B7/00—Compression machines, plants or systems, with cascade operation, i.e. with two or more circuits, the heat from the condenser of one circuit being absorbed by the evaporator of the next circuit
    • 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
    • F25B2500/18—Optimization, e.g. high integration of refrigeration components
    • 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
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2106—Temperatures of fresh outdoor air
    • 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
    • F25B49/00—Arrangement or mounting of control or safety devices
    • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
    • 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
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant the refrigerant being carbon dioxide

Definitions

  • the present invention relates to a two-stage refrigeration cycle system.
  • WO2012/0066763 discloses a two-stage refrigeration cycle according to the preamble of claim 1.
  • JP2012/097993 discloses a heat pump type air-warming device.
  • a two-stage refrigeration cycle system including a higher-stage refrigeration cycle that is a refrigeration cycle apparatus for circulating a higher-temperature refrigerant therein and a lower-stage refrigeration cycle that is a refrigeration cycle apparatus for circulating a lower-temperature refrigerant therein has been conventionally used as a system for cooling at a low temperature of several tens of minus degrees C.
  • a lower-stage refrigeration cycle and a higher-stage refrigeration cycle are coupled to each other by a cascade condenser configured to perform heat exchange between a lower-stage condenser in the lower-stage refrigeration cycle and a higher-stage evaporator in the higher-stage refrigeration cycle.
  • an auxiliary radiator is disposed at a stage previous to a cascade condenser in a lower-stage refrigeration cycle (see Patent Literature 1).
  • refrigerant discharged from a lower-temperature compressor is caused to reject heat by the auxiliary radiator to enhance the operating efficiency.
  • the two-stage refrigeration cycle system can be used in a range with a high efficiency in terms of a high compression ratio, this system is generally advantageous for energy conservation.
  • this system is generally advantageous for energy conservation.
  • even a single-stage refrigeration cycle system can easily perform cooling to a necessary temperature.
  • the use of a two-stage refrigeration cycle system does not have an advantage in energy conservation.
  • the operation mode is switched between the single-stage operation and a two-stage operation by connecting part of a higher-stage heat exchanger that is a radiator of the higher-stage refrigeration cycle in parallel to a condenser of the lower-stage refrigeration cycle so that the higher-stage heat exchanger can partially serve as a radiator of the lower-stage refrigeration cycle. That is, in the single-stage operation, the lower-stage refrigeration cycle transfers heat to outdoor air by using the part of the higher-stage heat exchanger.
  • the lower-stage refrigeration cycle prevents refrigerant from flowing into the part of the higher-stage heat exchanger and allows refrigerant to flow toward a cascade condenser, whereas the higher-stage refrigeration cycle transfers heat to outdoor air by using another part of the higher-stage heat exchanger.
  • the two-stage refrigeration cycle system of Patent Literature 1 is intended to enhance the operating efficiency by using the auxiliary radiator.
  • a low compression ratio such as an operation at a low outdoor-air temperature, for example, compressor performance inevitably degrades.
  • a temperature difference between a lower-stage condensing temperature and a higher-stage evaporating temperature in a cascade condenser leads to a loss.
  • this temperature difference does not cause any problem.
  • the influence of the temperature difference is larger than that in the operation at the high compression ratio, and thus, the operating efficiency inevitably degrades. That is, in the two-stage refrigeration cycle system of Patent Literature 1, although the use of the auxiliary radiator can enhance the operating efficiency, the operating efficiency inevitably degrades in a certain period, and it is difficult to obtain an energy conservation effect through the year.
  • Patent Literature 2 The lower-stage refrigeration cycle of the two-stage refrigeration cycle system described in Patent Literature 2 can avoid performance degradation to a certain degree by switching from the two-stage operation to the single-stage operation in an operation at a low compression ratio at a low outdoor-air temperature.
  • the cascade condenser rejects heat in the two-stage operation (where both the lower-stage refrigeration cycle and the higher-stage refrigeration cycle operate), whereas the higher-stage condenser rejects heat in the single-stage operation.
  • the channel needs to be switched between the two-stage operation and the single-stage operation, and part of the higher-stage condenser is wastefully unused in the two-stage operation. Accordingly, Patent Literature 2 has problems of an increase in costs due to addition of air-cooled radiators and switching valves arranged in parallel and addition of bypass channels.
  • the present invention has been made in view of the problems described above, and provides a two-stage refrigeration cycle system that can enhance operating efficiency by using an auxiliary radiator, avoid performance degradation in an operation at a low compression ratio, and obtain an energy conservation effect through the year without an increase in costs.
  • the present invention provides a two-stage refrigeration cycle system as set forth in claim 1.
  • the lower-stage refrigeration cycle includes the auxiliary radiator so that the auxiliary radiator is used as an aid for the lower-stage condenser in the two-stage operation and the auxiliary radiator is used as a main radiator in the single-stage operation, and an operation with a higher COP is selected from the two-stage operation and the single-stage operation.
  • the auxiliary radiator is used as a main radiator so that the lower-stage refrigeration cycle can be used without a change in channel.
  • Fig. 1 illustrates a configuration of a two-stage refrigeration cycle system according to an example useful in understanding the invention.
  • the two-stage refrigeration cycle system according to the example includes a lower-stage refrigeration cycle 10 and a higher-stage refrigeration cycle 20, and the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20 constitute refrigerant circuits in which refrigerant circulates, independently of each other.
  • a cascade condenser (inter-refrigerant heat exchanger) C is formed by coupling a higher-stage evaporator 24 and a lower-stage condenser 12 in such a manner that heat can be exchanged between refrigerant passing through the higher-stage evaporator 24 and refrigerant passing through the lower-stage condenser 12.
  • the two-stage refrigeration cycle system also includes a controller 30 that controls an operation of the entire two-stage refrigeration cycle system.
  • the levels of, for example, the temperature and pressure are not determined based on specific absolute values, and are determined relative to the states, operations, and other factors in, for example, a system or a device.
  • the lower-stage refrigeration cycle 10 constitutes a refrigerant circuit by sequentially connecting a lower-stage compressor 11, an auxiliary radiator 15, a lower-stage condenser 12, a lower-stage expansion valve (lower-stage expansion device) 13, and a lower-stage evaporator 14 by refrigerant pipes.
  • the higher-stage refrigeration cycle 20 constitutes a refrigerant circuit by sequentially connecting a higher-stage compressor 21, a higher-stage condenser 22, a higher-stage expansion valve (higher-stage expansion device) 23, and a higher-stage evaporator 24 by refrigerant pipes.
  • the lower-stage compressor 11 of the lower-stage refrigeration cycle 10 sucks refrigerant, compresses the refrigerant into a high-temperature high-pressure state, and discharges the resulting refrigerant.
  • This compressor is of a type whose rotation speed is controlled by, for example, an inverter circuit so that the discharging amount of higher-stage refrigerant can be adjusted.
  • the auxiliary radiator 15 serves as, for example, a gas cooler, and cools gas refrigerant discharged from the lower-stage compressor 11 through heat exchange with air in the outside (outdoor air), water, or brine, for example.
  • the auxiliary radiator 15 of Embodiment 1 performs heat exchange between outdoor air (ambient air) and refrigerant.
  • the lower-stage condenser 12 condenses refrigerant that has passed through the auxiliary radiator 15, through heat exchange with refrigerant that has passed from the higher-stage expansion valve 23 in the higher-stage refrigeration cycle 20, and obtains liquid refrigerant (condenses and liquefies the refrigerant).
  • a heat transfer tube through which refrigerant flowing in the lower-stage refrigeration cycle 10 passes serves as the lower-stage condenser 12.
  • the lower-stage expansion valve 13 serving as a pressure reducing device or an expansion device, for example, reduces the pressure of refrigerant flowing in the lower-stage refrigeration cycle 10 and expands the refrigerant.
  • the lower-stage expansion valve 13 is constituted by a flow rate controlling means such as an electronic expansion valve, a capillary tube (capillary), and a refrigerant flow rate adjusting means such as a themosensitive expansion valve.
  • the lower-stage evaporator 14 causes refrigerant flowing in the lower-stage refrigeration cycle 10 to exchange heat with, for example, a cooling target and evaporate as a gaseous refrigerant (evaporates and gasifies the refrigerant).
  • the cooling target is directly or indirectly cooled through heat exchange with refrigerant.
  • the higher-stage compressor 21 of the higher-stage refrigeration cycle 20 sucks refrigerant flowing in the higher-stage refrigeration cycle 20, compresses the refrigerant into a high-temperature high-pressure state, and discharges the refrigerant.
  • the higher-stage compressor 21 is, for example, a compressor of a type including, for example, an inverter circuit and capable of adjusting the discharging amount of refrigerant.
  • the higher-stage condenser 22 performs heat exchange between outdoor air, water, or brine, for example, and refrigerant flowing in the higher-stage refrigeration cycle 20, and condenses and liquefies the refrigerant.
  • the higher-stage condenser 22 performs heat exchange between outdoor air (ambient air) and refrigerant and incudes a higher-stage condenser fan 25 for promoting heat exchange.
  • the higher-stage condenser fan 25 is also a fan capable of adjusting an airflow rate.
  • the higher-stage expansion valve 23 serving as, for example, a pressure reducing device or an expansion device reduces the pressure of refrigerant flowing in the higher-stage refrigerant circuit and expands the refrigerant.
  • the higher-stage expansion valve 23 is a flow rate controlling means such as the electronic expansion valve described above or a refrigerant flow rate adjusting means such as a capillary tube.
  • the higher-stage evaporator 24 evaporates and gasifies refrigerant flowing in the higher-stage refrigeration cycle 20 through heat exchange.
  • a heat transfer tube through which refrigerant flowing in the higher-stage refrigeration cycle 20 passes in the cascade condenser C serves as the higher-stage evaporator 24 and performs heat exchange with refrigerant flowing in the lower-stage refrigeration cycle 10.
  • the cascade condenser C is an inter-refrigerant heat exchanger having the functions of the higher-stage evaporator 24 and the lower-stage condenser 12 and enables heat exchange between higher-stage refrigerant and lower-stage refrigerant.
  • the higher-stage refrigeration cycle 20 and the lower-stage refrigeration cycle 10 have a multi-stage configuration using the cascade condenser C and perform heat exchange between refrigerants, thereby causing these independent refrigerant circuits to cooperate with each other.
  • the controller 30 controls operations of devices constituting the two-stage refrigeration cycle system, for example.
  • An outdoor-air temperature detecting means 31 is a temperature sensor for detecting an outdoor-air temperature.
  • the outdoor-air temperature is a temperature used for detecting the outdoor-air temperature detecting means 31.
  • refrigerant used in the lower-stage refrigeration cycle 10 is CO 2 (carbon dioxide) because of the following reasons.
  • the lower-stage refrigeration cycle 10 is connected to, for example, a loaded device in a room, such as a showcase in a supermarket, and a refrigerant circuit is opened by rearranging the showcase so that leakage of refrigerant is likely to occur.
  • CO 2 carbon dioxide
  • refrigerant used in the higher-stage refrigeration cycle 20 is preferably a refrigerant whose influence on global warming is small, such as an HFO refrigerant (e.g., HFO1234yf, HFO1234ze), HC refrigerant, CO 2 , ammonia, and water.
  • HFO refrigerant e.g., HFO1234yf, HFO1234ze
  • HC refrigerant e.g., HFO1234yf, HFO1234ze
  • the higher-stage compressor 21 sucks higher-stage refrigerant, compresses the refrigerant into a high-temperature high-pressure state, and discharges the refrigerant.
  • the discharged refrigerant flows into the higher-stage condenser 22.
  • the higher-stage condenser 22 performs heat exchange between outdoor air supplied from the higher-stage condenser fan 25 and the higher-stage refrigerant, and condenses and liquefies the higher-stage refrigerant.
  • the pressure of the condensed and liquefied refrigerant is reduced by the higher-stage expansion valve 23.
  • the higher-stage refrigerant having a reduced pressure flows into the higher-stage evaporator 24 (cascade condenser C).
  • the higher-stage evaporator 24 performs heat exchange between the higher-stage refrigerant and lower-stage refrigerant passing through the lower-stage condenser 12, and evaporates and gasifies the higher-stage refrigerant.
  • the evaporated gaseous higher-stage refrigerant is sucked into the higher-stage compressor 21.
  • the lower-stage compressor 11 sucks CO 2 refrigerant, compresses the refrigerant into a high-temperature high-pressure state, and discharges the refrigerant.
  • the discharged refrigerant is cooled by the auxiliary radiator 15 and flows into the lower-stage condenser 12 (cascade condenser C).
  • the lower-stage condenser 12 performs heat exchange between the lower-stage refrigerant and the higher-stage refrigerant passing through the higher-stage evaporator 24 and condenses and liquefies the lower-stage refrigerant.
  • the pressure of the condensed and liquefied lower-stage refrigerant is reduced by the lower-stage expansion valve 13.
  • the lower-stage refrigerant having a reduced pressure flows into the lower-stage evaporator 14.
  • the lower-stage evaporator 14 evaporates and gasifies the lower-stage refrigerant through heat exchange with a cooling target.
  • the evaporated gaseous lower-stage refrigerant is sucked into the higher-stage compressor 21.
  • the higher-stage compressor 21 controls the frequency of a driving motor so as to control the cooling capacity of the higher-stage refrigeration cycle 20, thereby adjusting a pressure (high pressure) at a discharge side of the lower-stage refrigerant circuit. This will be specifically described.
  • Fig. 2 showing a relationship between an enthalpy and a saturation temperature in the two-stage refrigeration cycle system of the present invention.
  • a low pressure Ps and a high pressure Pd are fixed (that is, Fig. 2 shows a relationship between an enthalpy and a saturation temperature at a certain outdoor-air temperature).
  • a temperature difference ⁇ T arises between the lower-stage condensing temperature and the higher-stage evaporating temperature.
  • the temperature difference ⁇ T depends on the size (performance) of the cascade condenser C, and is, for example, about 5 degrees C.
  • the low pressure Ps refers to an evaporating pressure of the lower-stage refrigeration cycle 10 in each of a two-stage operation and a single-stage operation.
  • the high pressure Pd refers to a condensing pressure of the higher-stage refrigeration cycle 20 in the two-stage operation and to a condensing pressure of the lower-stage refrigeration cycle 10 in the single-stage operation.
  • the higher-stage evaporating temperature decreases, and the lower-stage condensing temperature (lower-stage high pressure) decreases accordingly.
  • the cooling capacity of the higher stage is reduced, the lower-stage high pressure increases.
  • the cooling load changes in accordance with the outdoor-air temperature, and a refrigeration capacity (corresponding to the evaporating capacity of the lower-stage refrigeration cycle 10) is determined with respect to the cooling load.
  • a refrigeration capacity (corresponding to the evaporating capacity of the lower-stage refrigeration cycle 10) is determined with respect to the cooling load.
  • a Gr refrigerant flow rate
  • the lower-stage compressor 11 is controlled in such a manner that the Gr (refrigerant flow rate) is constant if the ⁇ H (enthalpy difference) is constant.
  • CO 2 refrigerant used for the lower-stage refrigeration cycle 10 has a smaller refrigeration cycle effect than R32 used in the higher-stage refrigeration cycle 20.
  • R32 used in the higher-stage refrigeration cycle 20.
  • the volume of the higher-stage compressor 21 is increased so as to reduce the lower-stage high pressure, thereby reducing power consumption in the lower-stage refrigeration cycle 10.
  • a work of the higher-stage refrigeration cycle 20 is increased so that the operating efficiency of the entire two-stage refrigeration cycle system is enhanced.
  • CO 2 is often out of a supercritical state at the lower-stage high pressure of the lower-stage refrigeration cycle 10, and a saturation temperature (lower-stage condensing temperature) at which a phase change occurs in the lower-stage condenser 12 is fixed.
  • Fig. 3 shows a relationship between a lower-stage condensing temperature and a compressor input.
  • the abscissa represents a lower-stage condensing temperature
  • the ordinate represents a compressor input.
  • Fig. 3 shows an input of the higher-stage compressor 21, an input of the lower-stage compressor 11, and a total input thereof (a total input of the entire two-stage refrigeration cycle system). As shown in Fig.
  • the higher-stage compressor 21 of the two-stage refrigeration cycle system controls an operation in such a manner that the higher-stage compressor input and the lower-stage compressor input are substantially the same so as to maximize the COP.
  • the lower-stage condensing temperature of the lower-stage refrigeration cycle 10 is Tc
  • the total input is at the minimum
  • the COP is at the maximum.
  • the lower-stage refrigeration cycle 10 is controlled in such a manner that the lower-stage condensing temperature is kept at a target lower-stage condensing temperature Tc.
  • control is performed in such a manner that a temperature lower than the target lower-stage condensing temperature Tc by ⁇ T degrees C (which is 5 degrees C in this example as described above) is kept constant as a target higher-stage evaporating temperature.
  • ⁇ T degrees C which is 5 degrees C in this example as described above
  • Fig. 4 is a flowchart showing a flow before the rotation speed of the higher-stage compressor 21 is determined in accordance with the outdoor-air temperature in a two-stage operation of the two-stage refrigeration cycle system illustrated in Fig. 1 .
  • the refrigeration capacity is determined based on a request (a cooling load of use side equipment such as a cold room used by a user) (e.g., 10 kW for 10 horsepower).
  • the target lower-stage condensing temperature Tc at which the COP is at the maximum depends on a lower-stage evaporating temperature (e.g., -40 degrees C) determined based on a request from a user and a higher-stage condensing temperature uniquely determined in accordance with the outdoor-air temperature.
  • the higher-stage condensing temperature tends to increase as the outdoor-air temperature increases.
  • the target lower-stage condensing temperature Tc is determined based on the low pressure Ps converted from a lower-stage evaporating temperature ET and the high pressure Pd converted from a higher-stage condensing temperature CT in Fig. 2 .
  • the temperature difference ⁇ T is 5 degrees C and depends on the size (performance) of the cascade condenser C.
  • a necessary heat transfer amount Q1 required for the lower-stage condenser 12 in order to obtain this refrigeration capacity is determined.
  • the heat transfer amount in the lower-stage condenser 12 is equal to the heat absorption amount in the higher-stage evaporator 24.
  • a refrigeration capacity Q2 of the higher-stage refrigeration cycle 20 is determined.
  • the target lower-stage condensing temperature Tc is determined.
  • the controller 30 previously holds an approximation for obtaining the target lower-stage condensing temperature Tc using the lower-stage evaporating temperature and the higher-stage condensing temperature as variables and maximizing the COP and a corresponding map, and based on these information items, the target lower-stage condensing temperature Tc is determined. Since the higher-stage condensing temperature varies in accordance with the outdoor-air temperature, the target lower-stage condensing temperature Tc for maximizing the COP also varies in accordance with the outdoor-air temperature. Specifically, as the outdoor-air temperature increases, the higher-stage condensing temperature tends to increase, and the target lower-stage condensing temperature Tc also tends to increase.
  • a higher-stage evaporating temperature is determined by subtracting a temperature difference ⁇ T from the target lower-stage condensing temperature Tc.
  • an evaporating pressure of the higher-stage refrigeration cycle 20 is uniquely determined.
  • a refrigerant density ⁇ and an enthalpy difference ⁇ H of the higher-stage refrigeration cycle are determined.
  • Refrigeration capacity Q 2 of higher-stage refrigeration cycle 20 Gr refrigerant flow rate of higher-stage refrigeration cycle ⁇ ⁇ ⁇ H enthalpy difference in higher-stage cascade condenser
  • Refrigerant flow rate Gr of higher-stage refrigeration cycle 20 ⁇ refrigerant density of higher-stage refrigeration cycle 20 ⁇ Vst displacement amount of higher-stage compressor 21 ⁇ N rotation speed of higher-stage compressor 21 where although the "refrigeration capacity Q2 of higher-stage refrigeration cycle 20" and the "enthalpy difference ⁇ H" are already known, a "refrigerant flow rate Gr of higher-stage refrigeration cycle 20" is obtained from Equation (1).
  • the displacement amount Vst of the higher-stage compressor 21 is a unique value to the compressor and is already known.
  • the higher-stage compressor 21 is operated at the thus-determined rotation speed so that an operation can be controlled with substantially the same degree of the higher-stage compressor input and the lower-stage compressor input, thereby maximizing the COP.
  • the target higher-stage evaporating temperature is determined so as to control the higher-stage evaporating temperature.
  • the lower-stage condensing temperature may be directly detected for control.
  • the higher-stage compressor input and the lower-stage compressor input may also be directly detected or computed so as to control the higher-stage refrigeration cycle 20.
  • the higher-stage refrigeration cycle 20 may be controlled in accordance with the map or the approximation. As shown in Fig.
  • the control of the two-stage refrigeration cycle system is not limited to the control method of calculating the rotation speed of the higher-stage compressor 21 satisfying a target value based on principles of the refrigeration cycle, and may be a feedback control method based on a deviation between a target value (target lower-stage condensing temperature Tc) and a current value (current lower-stage condensing temperature) (a similar method can be applied to the lower-stage evaporating temperature).
  • a flow before determination of a "rotation speed of lower-stage compressor 11" is similar to the flow described above. That is, the determination is made in the order of a lower-stage evaporating temperature, a lower-stage evaporating pressure, a lower-stage refrigerant density and a lower-stage enthalpy difference, a lower-stage flow rate, and the rotation speed of the lower-stage compressor 11.
  • the lower-stage high pressure (lower-stage condensing temperature) is reduced in order to reduce power consumption of the lower-stage refrigeration cycle 10 having a low efficiency.
  • this description is only in control principle, and does not mean that the lower-stage high pressure is reduced in an actual operation.
  • the target lower-stage condensing temperature Tc is maintained at a constant value as described above.
  • R32 used in the higher-stage refrigeration cycle 20 is a refrigerant with a higher efficiency (refrigerant for obtaining a higher COP) than CO 2 refrigerant used in the lower-stage refrigeration cycle 10.
  • a slope ⁇ h on Mollier charts of Fig. 2 derived by an operation of the higher-stage compressor 21 is larger than a slope ⁇ l derived by an operation of the lower-stage compressor 11.
  • the higher-stage compressor input does not exceed the lower-stage compressor input until the lower-stage condensing temperature reaches the target lower-stage condensing temperature Tc. Then, at the target lower-stage condensing temperature Tc, the higher-stage compressor input becomes equal to the lower-stage compressor input.
  • the condensing temperature is 40 degrees C (where a supercritical CO 2 high pressure is 8.8 MPa), the degree of suction superheat is 5 degrees C, and the degree of liquid subcooling is 5 degrees C
  • theoretical COPs for individual refrigerants are CO 2 : 1.25, R32: 1.98, HFO1234yf: 1.84, HFO1234ze: 1.97, propane: 1.99, isobutane: 2.05, ammonia: 2.07, R134a: 2.01, R410A: 1.91, R407C: 1.98, and R404A: 1.76.
  • CO 2 has a lower COP and a lower efficiency than the other refrigerants such as HFO refrigerant, HFC refrigerant, and HC refrigerant.
  • CO 2 is used as refrigerant in the lower-stage refrigeration cycle 10.
  • the target lower-stage condensing temperature Tc is about 20 degrees C, and the target lower-stage condensing temperature Tc is lower than the outdoor-air temperature. Since the lower-stage compressor input in the lower-stage refrigeration cycle 10 having a low operating efficiency can be reduced by reducing the lower-stage high pressure (lower-stage condensing temperature) as described above, the target lower-stage condensing temperature Tc is located in a temperature range lower than the outdoor-air temperature.
  • the target lower-stage condensing temperature Tc is located in the temperature range lower than the outdoor-air temperature in a case where the CO 2 refrigerant having a low efficiency is applied to the lower-stage refrigeration cycle 10.
  • the locational relationship is not limited to this case depending on the combination of refrigerant types used in the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20.
  • the target lower-stage condensing temperature Tc is higher than the outdoor-air temperature in an operation at a low outdoor-air temperature
  • the target lower-stage condensing temperature Tc is lower than the outdoor-air temperature in an operation at a high outdoor-air temperature.
  • the correlation between the target lower-stage condensing temperature Tc and the outdoor-air temperature changes with respect to a change in the outdoor-air temperature with some combinations of refrigerants.
  • the amount of heat transfer of the auxiliary radiator 15 will be described.
  • the target lower-stage condensing temperature Tc is lower than the outdoor-air temperature at a high-temperature outdoor air condition of 32 degrees C.
  • the auxiliary radiator 15 transfers heat of lower-stage refrigerant to the outdoor air.
  • the temperature of the lower-stage refrigerant decreases only to the outdoor-air temperature at the lowest.
  • the amount of heat transfer differs even when the auxiliary radiator 15 reduces the discharge temperature of the lower-stage refrigerant to the same outdoor-air temperature.
  • Fig. 5 shows Mollier charts of heat transfer amounts in the case where the lower-stage condensing temperature is lower than the outdoor-air temperature and the case where the lower-stage condensing temperature is higher than the outdoor-air temperature.
  • Fig. 5(1) shows a heat transfer enthalpy difference in the case where the lower-stage condensing temperature is higher than the outdoor-air temperature
  • Fig. 5(2) shows a heat transfer enthalpy difference in the case where the lower-stage condensing temperature is lower than the outdoor-air temperature.
  • the temperature (temperature at point a) of refrigerant discharged from the lower-stage compressor 11 is, for example, 80 degrees C to 90 degrees C
  • the outdoor-air temperature is 20 degrees C
  • the lower-stage condensing temperature is 25 degrees C
  • the temperature (point a) of refrigerant at 80 degrees C to 90 degrees C first deceases to 25 degrees C (point b) that is a condensing temperature while remaining in a gaseous state, through heat exchange with the outdoor air in the auxiliary radiator 15 as shown in Fig. 5(1) .
  • the refrigerant is condensed into a liquid state (point c) while remaining at 25 degrees C.
  • the outdoor-air temperature is 20 degrees C
  • heat of the refrigerant can be further transferred, and the temperature of the refrigerant decreases to 20 degrees C (point d) in a liquid state.
  • the refrigerant is condensed when the condensing temperature is higher than the outdoor-air temperature, cooling with a phase change can be performed, and the amount of heat transfer is larger than that of cooling without a phase change.
  • the temperature (temperature at point a) of refrigerant discharged from the lower-stage compressor 11 is, for example, 80 degrees C to 90 degrees C
  • the outdoor-air temperature is 20 degrees C
  • the lower-stage condensing temperature is 10 degrees C
  • the auxiliary radiator 15 transfers heat to the outdoor air
  • the temperature of refrigerant at 80 degrees C to 90 degrees C decreases only to 20 degrees C that is the outdoor-air temperature to the lowest, through heat exchange with the outdoor air in the auxiliary radiator 15.
  • the temperature (point a) of refrigerant at 80 degrees C to 90 degrees C decreases to 20 degrees C (point b) in the auxiliary radiator 15 while remaining in a gaseous state.
  • the auxiliary radiator 15 cannot perform cooling with a phase change but performs gas cooling without a phase change. That is, the auxiliary radiator 15 is used in a gas cooling range.
  • the amount of heat transfer by the auxiliary radiator 15 cannot be increased and is, at most, increased to the amount of heat transferred until the discharged refrigerant has its temperature decrease to the outdoor-air temperature while maintaining a gaseous state.
  • Fig. 6 is a diagram for showing a relationship between the amount of heat transfer by the auxiliary radiator 15 and a COP.
  • Fig. 6 is a Mollier chart of the lower-stage refrigeration cycle 10.
  • a comparison between a case where the amount of heat transfer by the auxiliary radiator 15 is Qsub1 in Fig. 6 and a case where the amount of heat transfer by the auxiliary radiator 15 is Qsub2 in Fig. 6 shows that a heat transfer amount Qc2 ( ⁇ Qc1) of the corresponding lower-stage condenser 12 in the case of Qsub2 is smaller than that in the other case.
  • the amount of heat exchange of the higher-stage evaporator 24 is equal to that of the lower-stage condenser 12.
  • the higher-stage refrigeration cycle 20 only needs to obtain a balance with the heat transfer amount Qc2 in the lower-stage condenser 12, and thus, the higher-stage compressor input in the case where the heat transfer amount of the auxiliary radiator 15 is Qsub2 than that in the case where the heat transfer amount of the auxiliary radiator 15 is Qsub1.
  • the COP value increases.
  • the auxiliary radiator 15 is used in the gas cooling range.
  • the temperature of discharged refrigerant is reduced to the outdoor-air temperature even at the maximum heat transfer.
  • the COP can be increased.
  • the heat transfer amount of the auxiliary radiator 15 is obtained to such a degree that the discharge temperature of refrigerant at the auxiliary radiator 15 can be reduced to a temperature around the outdoor-air temperature.
  • the "heat transfer amount of the auxiliary radiator 15 in reducing the discharge temperature of refrigerant to a temperature around the outdoor-air temperature through heat transfer in the auxiliary radiator 15" will be hereinafter referred to as a necessary heat transfer amount.
  • the airflow rate of the auxiliary radiator 15 is controlled or the configuration of the auxiliary radiator 15 itself is designed, for example. In this manner, the heat transfer amount of the auxiliary radiator 15 is set at the necessary heat transfer amount, thereby increasing the COP as compared to the case of a heat transfer amount smaller than the necessary heat transfer amount.
  • the necessary heat transfer amount varies in accordance with the outdoor-air temperature.
  • a necessary heat transfer amount in a low-temperature outdoor air condition and a necessary heat transfer amount in a high-temperature outdoor air condition need to be known.
  • the auxiliary radiator 15 is used in the gas cooling range and the necessary heat transfer amount is smaller as described above.
  • the target lower-stage condensing temperature Tc is higher in an operation at a low outdoor-air temperature and the target lower-stage condensing temperature Tc is lower in an operation at a high outdoor-air temperature, for example.
  • the correlation of the target lower-stage condensing temperature Tc with the outdoor-air temperature changes, and the necessary heat transfer amount changes.
  • the auxiliary radiator 15 cannot perform cooling with a phase change, and the necessary heat transfer amount decreases, as described above.
  • the heat transfer amount of the auxiliary radiator 15 is, at most, the amount of heat transferred until the discharge refrigerant has its temperature reduced to the outdoor-air temperature while remaining in the gaseous state.
  • the heat transfer amount of the auxiliary radiator 15 cannot increase.
  • the airflow rate of the auxiliary radiator 15 is reduced and optimized, fan input is unnecessarily consumed, leading to a COP decrease.
  • the heat transfer amount of the auxiliary radiator 15 is obtained to such a degree that the auxiliary radiator 15 can reduce the discharge temperature of refrigerant to a temperature around the outdoor-air temperature without wastefully increasing fan input.
  • fan input can be optimized, and the COP of the entire two-stage refrigeration cycle system can be enhanced.
  • the auxiliary radiator 15 performs cooling with a phase change, and the necessary heat transfer amount increases. At this time, the airflow rate of the auxiliary radiator 15 is continuously increased as the necessary heat transfer amount increases, and the COP of the entire two-stage refrigeration cycle system can be enhanced by increasing the heat transfer amount of the auxiliary radiator 15.
  • the airflow rate control of the auxiliary radiator 15 with respect to a change in the necessary heat transfer amount is performed in the following manner. Specifically, the control is performed in such a manner that a temperature difference between an outlet refrigerant temperature of the auxiliary radiator 15 and the outdoor-air temperature is a predetermined value (which is about 2 degrees C in this example). In this manner, the airflow rate of the auxiliary radiator 15 can be appropriately adjusted, thereby enhancing the COP of the entire two-stage refrigeration cycle system.
  • the heat transfer amount of the auxiliary radiator 15 with respect to the outdoor-air temperature can be appropriately controlled by controlling the airflow rate of the auxiliary radiator 15 so that a high degree of energy conservation can be obtained through the year.
  • a sufficient heat transfer area of the auxiliary radiator 15 is about 10 to 20% of the heat transfer area of the higher-stage condenser 22, and is sufficient.
  • the heat transfer area of the auxiliary radiator 15 is enlarged to a degree substantially equal to that of the higher-stage condenser 22, and the COP of the entire two-stage refrigeration cycle system can be enhanced by significantly increasing the heat transfer amount of the auxiliary radiator 15.
  • the auxiliary radiator 15 and the higher-stage condenser 22 can have similar shapes so that components can be commonly used, thereby enabling cost reduction.
  • the airflow rate of the auxiliary radiator 15 is increased as the necessary heat transfer amount increases, and the heat transfer amount of the auxiliary radiator 15 significantly increases.
  • the heat transfer amount of the auxiliary radiator 15 significantly increases, the heat transfer amount in the lower-stage condenser 12 of the cascade condenser C decreases, and the cooling capacity at the higher-stage refrigeration cycle also decreases.
  • the lower-stage condensing temperature cannot be controlled by promoting heat transfer in the lower-stage condenser 12 through control of the cooling capacity of the higher-stage refrigeration cycle. That is, when the heat transfer amount of the auxiliary radiator 15 greatly exceeds the heat transfer amount of the lower-stage condenser 12, the lower-stage condensing temperature depends on the heat transfer amount of the auxiliary radiator 15.
  • the lower-stage condensing temperature is higher than the outdoor-air temperature by a predetermined temperature (about 10 degrees C in this example) so that the lower-stage compressor input and the fan input can be optimized and the COP of the entire two-stage refrigeration cycle system can be enhanced.
  • the auxiliary radiator 15 provided in the lower-stage refrigeration cycle 10 is used as an aid for the lower-stage condenser 12 so as to enhance the operating efficiency in a two-stage operation, whereas the auxiliary radiator 15 is used as a main radiator in a single-stage operation. That is, a feature of the present invention is that the auxiliary radiator 15 is used in both the two-stage operation and the single-stage operation. In an operation at a low outdoor-air temperature, the operation mode is switched to the single-stage operation whose COP is higher than that of the two-stage operation. In this manner, as another feature of the present invention, an operation with a higher COP is selected from the two-stage operation and the single-stage operation. In addition, as still another feature of the present invention, in performing the single-stage operation, the auxiliary radiator 15 is used as a main radiator so that the lower-stage refrigeration cycle 10 can be used without switching the channel, unlike Patent Literature 2.
  • the operating efficiency can be enhanced by using the auxiliary radiator 15, and in addition, performance degradation in an operation at a low compression ratio is avoided so that an energy conservation effect can be achieved through the year.
  • the two-stage refrigeration cycle system When the two-stage refrigeration cycle system operates at a low outdoor-air temperature, the higher-stage condensing temperature decreases, and the lower-stage condensing temperature decreases accordingly.
  • each of the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20 operates at a low compression ratio, the compressor performance thereof degrades, and operations of these cycles deviate from operation ranges defined as standards. Thus, reliability cannot be maintained.
  • the influence of performance degradation increases.
  • compressor performance changes depending on the compression ratio and the rotation speed.
  • a compressor is designed such that the compressor performance is at the maximum at a compression ratio at which the compressor is supposed to be most frequently used.
  • performance degrades significantly.
  • the two-stage refrigeration cycle system of the Embodiment 1 performs the following control. Specifically, if it is determined that the outdoor-air temperature has decreased to the "predetermined outdoor-air temperature" or less, which might cause a COP decrease if the two-stage operation continues without change, in the two-stage operation, the controller 30 stops the higher-stage refrigeration cycle 20 and causes only the lower-stage refrigeration cycle 10 to operate. That is, the two-stage operation is switched to the single-stage operation.
  • a specific flow of refrigerant in the single-stage operation is as follows.
  • refrigerant compressed in the lower-stage compressor 11 and discharged therefrom is subjected to heat transfer and cooling only by using the auxiliary radiator 15, and the pressure of the refrigerant cooled by the auxiliary radiator 15 is reduced by the lower-stage expansion valve 13. Then, the resulting refrigerant is evaporated by the lower-stage evaporator 14 and is fed back to the lower-stage compressor 11.
  • the "predetermined outdoor-air temperature,” that is, a threshold outdoor-air temperature used for determining whether a COP decrease is induced or not will be described later.
  • the two-stage operation is switched to the single-stage operation, and only the lower-stage refrigeration cycle 10 operates.
  • the lower-stage compressor 11 is kept at an appropriate compression ratio so that sufficient performance and reliability can be obtained.
  • structural problems typically found in the two-stage refrigeration cycle system can be avoided at the same time. Specifically, it is possible to avoid performance degradation caused by an increased ratio of the temperature difference ⁇ T of the cascade condenser C with respect to the temperature difference between the lower-stage evaporating temperature and the higher-stage condensing temperature.
  • the auxiliary radiator 15 is used as a main radiator without change including switching of channel.
  • a threshold outdoor-air temperature as a switching point between the two-stage operation and the single-stage operation will be described.
  • Fig. 7 shows an outdoor-air temperature to COP characteristic in each of a single-stage operation and a two-stage operation (with an auxiliary radiator) in the two-stage refrigeration cycle system illustrated in Fig. 1 .
  • Fig. 7 also shows an outdoor-air temperature to COP characteristic (two-stage operation (without an auxiliary radiator)) in a two-stage operation of a conventional two-stage refrigeration cycle system including no auxiliary radiator 15 for comparison.
  • Fig. 7 is a graph showing outdoor-air temperature to COP characteristics in a case where a low pressure Ps is fixed at a certain pressure value and a high pressure Pd is determined in accordance with the outdoor-air temperature.
  • a temperature Tca at a point at which a characteristic of the single-stage operation and a characteristic of the two-stage operation (with an auxiliary heat exchanger) intersect each other is a threshold temperature for determining whether a COP decrease occurs or not. That is, if the outdoor-air temperature is higher than the threshold outdoor-air temperature Tca, the COP in the two-stage operation is higher than that in the single-stage operation, whereas if the outdoor-air temperature is less than or equal to the threshold outdoor-air temperature Tca, the COP of the single-stage operation is higher than that of the two-stage operation.
  • an operation is performed with switching the operation mode to an operation with a higher COP based on a comparison result between the outdoor-air temperature and the threshold outdoor-air temperature Tca.
  • the characteristics shown in Fig. 7 are obtained beforehand by experiment or simulation, and thus, the threshold outdoor-air temperature Tca can also be obtained beforehand.
  • Tc is a target lower-stage condensing temperature in the two-stage operation (with an auxiliary radiator).
  • the rate of a COP increase in the two-stage operation becomes high from the Tc. This is because if the outdoor-air temperature is less than or equal to the lower-stage condensing temperature Tc as described with reference to Fig. 5 , lower-stage refrigerant is condensed in the auxiliary radiator 15 so that cooling with a phase change can be performed, and the amount of heat transfer is larger than that in cooling without a phase change.
  • the COP increases.
  • a comparison between a case where the outdoor-air temperature is less than or equal to the lower-stage condensing temperature Tc and a case where the outdoor-air temperature is higher than the lower-stage condensing temperature Tc shows that the rate of the COP increase is higher in the case where the outdoor-air temperature is less than or equal to the lower-stage condensing temperature Tc than that in the case where the outdoor-air temperature is higher than the lower-stage condensing temperature Tc.
  • switching from the single-stage operation to the two-stage operation is also based on the threshold outdoor-air temperature Tca, and an operation mode with a high COP is selected.
  • the switching direction from the two-stage operation to the single-stage operation and switching in the reverse direction are provided with a hysteresis or switching is not performed in a predetermined time. Since the threshold outdoor-air temperature Tca changes with the lower-stage evaporating temperature, the threshold outdoor-air temperature Tca may be set by using an approximation using the lower-stage evaporating temperature as a variable or a map.
  • CO 2 refrigerant is used in the lower-stage refrigeration cycle 10 as described above.
  • the threshold outdoor-air temperature Tca used for switching from the two-stage operation to the single-stage operation in the case where CO 2 refrigerant is used in the lower-stage refrigeration cycle 10 is set in such a manner that the high pressure does not exceed the critical pressure after switching to the single-stage operation, which will be described below.
  • the threshold outdoor-air temperature Tca for use in switching from the two-stage operation to the single-stage operation in the case where CO 2 refrigerant is used in the lower-stage refrigeration cycle 10 is 21 degrees C, which is lower by 10 degrees C than 31 degrees C that is a critical saturation temperature. That is, when the outdoor-air temperature decreases to 21 degrees C or less in the two-stage operation, the operation mode is switched to the single-stage operation.
  • the timing of returning from the single-stage operation to the two-stage operation is the time when a refrigerant temperature at the higher-pressure side becomes a critical saturation temperature of 31 degrees C.
  • Fig. 8 is a flowchart showing an operation of the two-stage refrigeration cycle system illustrated in Fig. 1 .
  • the process shown in the flowchart of Fig. 8 is repeatedly performed at every control interval, for example.
  • the controller 30 compares an outdoor-air temperature detected by the outdoor-air temperature detecting means 31 and a previously determined threshold outdoor-air temperature Tca (S11). If the outdoor-air temperature is less than or equal to the threshold outdoor-air temperature Tca, the controller 30 performs a single-stage operation (S12), whereas if the outdoor-air temperature is higher than the threshold outdoor-air temperature Tca, the controller 30 performs a two-stage operation (S13). Thus, if the outdoor-air temperature decreases to the threshold outdoor-air temperature Tca or less in the two-stage operation, the two-stage operation is switched to the single-stage operation. If the outdoor-air temperature exceeds the threshold outdoor-air temperature Tca in the single-stage operation, the single-stage operation is switched to the two-stage operation.
  • the outdoor-air temperature is used as a threshold value for switching from the two-stage operation to the single-stage operation.
  • the switching may be based on the following threshold values.
  • Compressor performance generally depends on a compression ratio, and thus, an operation switching may be performed based on the compression ratio of the lower-stage refrigeration cycle 10 or the higher-stage refrigeration cycle 20. Specifically, if the compression ratio of the lower-stage refrigeration cycle 10 or the higher-stage refrigeration cycle 20 is less than or equal to a threshold compression ratio that is a threshold value for switching, the single-stage operation is performed, whereas if the compression ratio of the lower-stage refrigeration cycle 10 or the higher-stage refrigeration cycle 20 is higher than the threshold compression ratio, the two-stage operation is performed.
  • the threshold compression ratio is set (at, for example, a compression ratio of 2.0) based on compressor performance or reliability.
  • the two-stage operation is switched to the single-stage operation.
  • the threshold compression ratio In this manner, it is possible to ensure prevention of compressor performance degradation caused by a high compression ratio.
  • the operation switching may be performed based on a compression ratio (compression ratio based on an evaporating pressure of the lower-stage refrigeration cycle 10 and a condensing pressure of the higher-stage refrigeration cycle 20) of the entire two-stage refrigeration cycle in the two-stage operation.
  • a compression ratio compression ratio based on an evaporating pressure of the lower-stage refrigeration cycle 10 and a condensing pressure of the higher-stage refrigeration cycle 20
  • the threshold compression ratio e.g., a compression ratio of 4.0
  • the target lower-stage condensing temperature Tc at which the COP is at the maximum is lower than the outdoor-air temperature, and as the outdoor-air temperature decreases, the target lower-stage condensing temperature Tc also decreases while maintaining this relationship. That is, as the outdoor-air temperature decreases, the target lower-stage condensing temperature Tc tends to decrease while being kept lower than the outdoor-air temperature. As the outdoor-air temperature decreases, the temperature difference between the outdoor-air temperature and the target lower-stage condensing temperature Tc decreases.
  • the two-stage operation is performed, whereas when the temperature difference between the outdoor-air temperature and the target lower-stage condensing temperature Tc becomes less than or equal to the predetermined value a, the two-stage operation may be switched to the single-stage operation. This will be described with reference to Fig. 9 .
  • Fig. 9 is a graph showing a relationship between an outdoor-air temperature and a threshold outdoor-air temperature Tca in the two-stage refrigeration cycle system according to the example, where the abscissa represents the outdoor-air temperature and the ordinate represents a temperature (outdoor-air temperature - Tc) obtained by subtracting the target lower-stage condensing temperature Tc from the outdoor-air temperature.
  • the threshold outdoor-air temperature Tca may be replaced by the predetermined value a obtained by subtracting the target lower-stage condensing temperature Tc from the outdoor-air temperature.
  • the operation mode may be switched in the following manner. Specifically, if the temperature obtained by subtracting the target lower-stage condensing temperature Tc from the outdoor-air temperature is higher than the predetermined value a, the two-stage operation may be performed, whereas if the temperature obtained by subtracting the target lower-stage condensing temperature Tc from the outdoor-air temperature is less than or equal to the predetermined value a, the single-stage operation may be performed.
  • Fig. 9 shows a temperature range in which the "outdoor-air temperature - Tc" decreases to a negative value, that is, Tc becomes higher than the outdoor-air temperature. This is because of the following reasons. A system has a lower limit of the compression ratio. Thus, even when the outdoor-air temperature decreases, the target lower-stage condensing temperature Tc does not decrease below a certain temperature. Accordingly, the relationship between the outdoor-air temperature and the Tc is reversed so that the Tc becomes higher than the outdoor-air temperature in some cases.
  • Switching from the single-stage operation to the two-stage operation is performed in the following manner. Specifically, when a high pressure Pd in the single-stage operation exceeds a high pressure Pd immediately after switching from the two-stage operation to the single-stage operation, the single-stage operation is switched to the two-stage operation. In this manner, it is possible to ensure selection of an operation mode with a high COP.
  • the threshold value used in switching from the single-stage operation to the two-stage operation in the forward direction (from the two-stage operation to the single-stage operation) and the threshold value used in switching from the two-stage operation to the single-stage operation in the reverse direction (from the single-stage operation to the two-stage operation) are not limited to the same threshold value, but may be different from each other. That is, the threshold outdoor-air temperature Tca in (1) may be used for switching in the forward direction whereas the high pressure Pd in (3) being used for switching in the reverse direction, for example.
  • the higher-stage condenser 22 and the auxiliary radiator 15 are plate-fin-tube type heat exchangers in which heat transfer tubes penetrate heat transfer fins having flat-plate shapes. As shown in Fig. 10 , the higher-stage condenser 22 and the auxiliary radiator 15 may share heat transfer fins 40 so as to form an integrated radiator 42, or heat transfer fin parts may be divided. If the heat transfer fins 40 are integrated, fabrication is easily performed in terms of the configuration of the heat exchangers.
  • Fig. 10 illustrates a configuration of the two-stage refrigeration cycle system in a case where the higher-stage condenser and the auxiliary radiator illustrated in Fig. 1 form an integrated radiator.
  • reference numeral 43 denotes an air-sending device for sending air to the integrated radiator 42 obtained by integrating the higher-stage condenser 22 and the auxiliary radiator 15.
  • the auxiliary radiator 15 Since a high-temperature refrigerant gas discharged from the lower-stage compressor 11 passes through the auxiliary radiator 15, the auxiliary radiator 15 is at a high temperature. Thus, in a case where the heat transfer fins are divided between the auxiliary radiator 15 at the high temperature and the higher-stage condenser 22, an excellent heat insulating effect is obtained, and both the auxiliary radiator 15 and the higher-stage condenser 22 can more efficiently transfer heat.
  • the auxiliary radiator 15 is disposed in an upper portion (above in the direction of gravity) and the higher-stage condenser 22 is disposed in a lower portion (below in the direction of gravity) as shown in Fig. 10 . Since the auxiliary radiator 15 at a high temperature is disposed in the upper portion of the heat exchangers, the heat transfer of the auxiliary radiator 15 does not interfere with the higher-stage condenser 22. That is, heat transfer target fluid heated by the auxiliary radiator 15 does not move toward the higher-stage condenser 22, and both the auxiliary radiator 15 and the higher-stage condenser 22 can efficiently transfer heat.
  • the integrated radiator 42 may be configured as illustrated in Fig. 11 .
  • Fig. 11 is an illustration of an example configuration in the case where the higher-stage condenser and the auxiliary radiator shown in Fig. 1 form the integrated radiator.
  • Fig. 11 schematically illustrates a configuration of the heat transfer fins.
  • the integrated radiator 42 includes a plurality of heat transfer fins 40 spaced from each other so as to allow air to pass therebetween, and a plurality of heat transfer tubes 41 penetrating the heat transfer fins 40.
  • the heat transfer tubes 41 are arranged in a plurality of columns in a direction perpendicularly to an air passage direction (vertically in Fig. 11 ) and in a plurality of rows in the air passage direction (laterally in Fig. 11 ).
  • the heat transfer tubes 41 constituting the auxiliary radiator 15 are collected in one row. This configuration can obtain a flow rate of the auxiliary radiator 15 with no waste in the single-stage operation. As a result, a high flow rate can be obtained in the single-stage operation, thereby enhancing performance.
  • the heat transfer area of the auxiliary radiator 15 is about 10 to 20% of the heat transfer area of the higher-stage condenser 22 in order to maximize a COP in a two-stage operation in a region where the auxiliary radiator 15 is frequently used in a gas range and the outdoor-air temperature is high through the year.
  • the heat transfer area of the auxiliary radiator 15 is preferably increased to be substantially equal to that of the higher-stage condenser 22 so as to enhance a COP in a single-stage operation at a low outdoor-air temperature.
  • components can be commonly used by setting the heat transfer areas of the auxiliary radiator 15 and the higher-stage condenser 22 at substantially the same degree, thereby enabling cost reduction.
  • the target lower-stage condensing temperature Tc at which the COP is at the maximum as a target value is used as a target value in such a manner that the compression ratio in the case of using a refrigerant with a lower theoretical COP is lower and the compression ratio in the case of using a refrigerant with a higher theoretical COP is higher. Since the compression ratios differ between the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20, the compression ratio might be extremely low or high. In particular, in the case of using CO 2 refrigerant with a low theoretical COP or mixed refrigerant including CO 2 in one of the refrigeration cycles, difference in the compression ratio becomes conspicuous.
  • the compression ratios of both the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20 do not become appropriate at the same time, at least one of these cycles shows degradation of compressor performance.
  • the single-stage operation capable of avoiding performance degradation at a low compression ratio or a high compression ratio is especially effective, thereby significantly enhancing an energy conservation effect through the year.
  • the lower-stage condensing temperature with which the compression ratio of the lower-stage refrigeration cycle is low is used as a target.
  • the low-pressure side compression ratio significantly decreases at the low outdoor-air temperature.
  • the single-stage operation of the lower-stage refrigeration cycle 10 that avoids performance degradation due to a compression ratio decrease of the lower-stage refrigeration cycle is especially effective, and an energy conservation effect through the year can be significantly enhanced.
  • refrigerant with a high theoretical COP examples include R32, R410A, R134a, R404A, R407C, HFO1234yf, HFO1234ze, ammonia, propane, and isobutane.
  • the present invention includes a configuration in which the refrigerant with a high theoretical COP is used in at least one of the lower-stage refrigeration cycle 10 and the higher-stage refrigeration cycle 20.
  • the lower-stage refrigeration cycle 10 includes the auxiliary radiator 15 so that the auxiliary radiator 15 is used as an aid for the lower-stage condenser 12 in the two-stage operation, whereas the auxiliary radiator 15 is used as a main radiator in the single-stage operation. Switching between the two-stage operation and the single-stage operation is performed in such a manner than an operation with a higher COP is performed. In this manner, a high operating efficiency can be obtained through the year, thereby achieving energy conservation. Since the auxiliary radiator 15 is used as a main radiator in the single-stage operation, the lower-stage refrigeration cycle 10 can be used without change including switching of channel. Thus, none of an air-cooled radiator, a switching valve, and a bypass channel needs to be additionally provided for the single-stage operation, and thus, cost reduction can be achieved.
  • CO 2 is used in a lower-stage refrigeration cycle 10 and a design pressure of the lower-stage refrigeration cycle 10 is reduced to about 4.15 MPa, which is a design pressure approximately equal to that of HFC refrigerant, corresponding to R410A, for example.
  • a refrigerant operating pressure of CO 2 is higher than that of a conventional HFC refrigerant such as R404A or R410A.
  • a conventional HFC refrigerant such as R404A or R410A.
  • an additional component with a high design pressure needs to be used, which leads to a significant cost increase.
  • component of the conventional lower-stage refrigeration cycle using an operating refrigerant as the HFC refrigerant are diverted.
  • the design pressure of the lower-stage refrigeration cycle 10 is not increased but is reduced to about 4.15 MPa, which is a design pressure approximately equal to that of the HFC refrigerant, corresponding to R410A, for example.
  • a configuration enabling the design pressure of the lower-stage refrigeration cycle to be reduced to about 4.15 MPa will be described with reference to Figs. 12 and 13 later.
  • the lower-stage compressor 11 Since a single-stage operation is performed in an operation with a low load, the lower-stage compressor 11 is supposed to be repeatedly started and stopped in the single-stage operation.
  • a lower-stage compressor 11 of the lower-stage refrigeration cycle 10 stops, refrigerant is heated to a temperature around an outdoor-air temperature and becomes gas so that the pressure in the lower-stage refrigeration cycle 10 increases.
  • the pressure in the lower-stage refrigeration cycle 10 might exceeds the design pressure with some inner capacity and amount of enclosed refrigerant in the lower-stage refrigeration cycle 10.
  • Such a pressure increase while the lower-stage refrigeration cycle 10 stops can be solved by starting a higher-stage refrigeration cycle 20 so as to cool the lower-stage refrigeration cycle 10.
  • alternate starting and stopping of the higher-stage refrigeration cycle 20 and the lower-stage refrigeration cycle 10 do not lead to energy conservation because of the influence of an ON/OFF loss.
  • the system cannot cope with a blackout and an abnormal situation such as a failure in the higher-stage compressor 21.
  • the two-stage refrigeration cycle system is configured as illustrated in Fig. 12 .
  • Fig. 12 illustrates an example configuration 1 of the two-stage refrigeration cycle system according to Embodiment 2 of the present invention.
  • the same reference characters designate the same parts in Fig. 1 .
  • an expansion tank 32 is connected to a line between the lower-stage compressor 11 and the lower-stage evaporator 14 of the lower-stage refrigeration cycle 10 illustrated in Fig. 1 through a solenoid valve 33.
  • the inner capacity of the lower-stage refrigeration cycle 10 can be enlarged by opening the solenoid valve 33 so that the expansion tank 32 communicates with the lower-stage refrigeration cycle 10.
  • the solenoid valve 33 is opened so that refrigerant in the lower-stage refrigeration cycle 10 is recovered to the expansion tank 32.
  • the solenoid valve 33 is of an exciting close type, and is open even in a blackout so that refrigerant can be recovered to the expansion tank 32. In this manner, since the refrigerant in the lower-stage refrigeration cycle 10 can be recovered to the expansion tank 32, it is possible to prevent the pressure in the lower-stage refrigeration cycle 10 from exceeding a set pressure.
  • the expansion tank 32 is provided in the low-pressure side, especially a suction port of the lower-stage compressor 11, in such a manner that refrigerant in the expansion tank 32 can be recovered to the lower-stage refrigeration cycle 10 when the lower-stage refrigeration cycle 10 is restarted.
  • the inside of the expansion tank 32 is always kept at a low pressure. If the expansion tank 32 has been cooled, refrigerant recovery from the lower-stage refrigeration cycle 10 to the expansion tank 32 can be further promoted.
  • FIG. 13 Another configuration as illustrated in Fig. 13 may also be employed.
  • Fig. 13 illustrates an example configuration 2 of the two-stage refrigeration cycle system according to the Embodiment of the present invention.
  • the same reference characters designate the same parts in Fig. 1 .
  • a second expansion valve (second expansion device) 34 is provided in an upstream part of a liquid pipe 16 between a cascade condenser C and a lower-stage expansion valve 13 in the lower-stage refrigeration cycle 10. Two-phase gas-liquid flows in the liquid pipe 16 so that the amount of refrigerant in the liquid pipe 16 is reduced. In this manner, the amount of enclosed refrigerant in the lower-stage refrigeration cycle 10 is reduced, and even when CO 2 becomes a critical state, it is possible to prevent the pressure in the lower-stage refrigeration cycle 10 from exceeding the design pressure.
  • Fig. 13 shows the configuration including no expansion tank 32, the expansion tank 32 may be provided. In this case, the use of the second expansion valve 34 can reduce the volume of the expansion tank 32 as compared to the configuration illustrated in Fig. 12 , and the size of the expansion tank 32 can be reduced.
  • Fig. 14 shows a relationship between an enthalpy and a saturation temperature in the two-stage refrigeration cycle system according to the Embodiment of the present invention.
  • each of the two-stage operation and the single-stage operation is performed in such a manner that the lower-stage condensing temperature is less than or equal to 8 degrees C, which is a CO 2 saturation temperature corresponding to a design pressure of 4.15 MPa.
  • the radiator is designed in such a manner that the lower-stage condensing temperature is higher than the outdoor-air temperature by about 10 degrees C as described above, switching from the two-stage operation to the single-stage operation is performed at -2 degrees C, which is an outdoor-air temperature lower than the lower-stage condensing temperature by 10 degrees C. That is, if the outdoor-air temperature is less than or equal to -2 degrees C, the single-stage operation is performed, and if the outdoor-air temperature is higher than -2 degrees C, the two-stage operation is performed.
  • a target value (lower-stage condensing temperature) is determined in the following manner. Specifically, in a case where the target value (lower-stage condensing temperature) with a maximum COP is less than or equal to 8 degrees C, an operation is performed using this target value, whereas in a case where the target value (lower-stage condensing temperature) with a maximum COP is higher than 8 degrees C, an operation is performed with the target value being controlled to 8 degrees.
  • the configuration in which the expansion tank 32 is provided in the lower-stage refrigeration cycle 10 or the configuration in which the amount of enclosed refrigerant is reduced by flowing refrigerant in a two-phase state in the liquid pipe with the second expansion valve 34 achieves the following advantages. Specifically, even when the higher-stage refrigeration cycle 20 stops, the pressure in the lower-stage refrigeration cycle 10 using CO 2 can be reduced to be less than or equal to a design pressure of 4.15 MPa corresponding to that of HFC refrigerant so that components of a conventional HFC refrigeration cycle system can be used. Thus, costs can be reduced.
  • the higher-stage refrigeration cycle 20 can always stop so that a loss due to ON/OFF can be avoided, thereby achieving an energy conservation effect.
  • the two-stage refrigeration cycle systems according to the Embodiments are widely applicable to cold storages or refrigeration cycle systems such as showcases, refrigerator-freezers for commercial use, and vending machines that need CFC elimination of refrigerant, reduction of CFC refrigerant, and energy conservation of the systems.

Landscapes

  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Air Conditioning Control Device (AREA)
  • Devices That Are Associated With Refrigeration Equipment (AREA)

Claims (16)

  1. Système de cycle de réfrigération à deux étages comprenant :
    un compresseur d'étage supérieur (21), un condenseur d'étage supérieur (22), un détendeur d'étage supérieur (23) et un évaporateur d'étage supérieur (24) connectés dans un cycle de réfrigération d'étage supérieur (20) par des tuyaux à travers lequel du fluide frigorigène circule ;
    un compresseur d'étage inférieur (11), un radiateur auxiliaire (15), un condenseur d'étage inférieur (12), un détendeur d'étage inférieur (13) et un évaporateur d'étage inférieur (14) connectés en série dans un cycle de réfrigération d'étage inférieur (10) par des tuyaux à travers lesquels du fluide frigorigène circule ;
    un condenseur en cascade (C) comprenant l'évaporateur d'étage supérieur (24) et le condenseur d'étage inférieur (12) et configuré pour échanger de la chaleur entre le fluide frigorigène circulant dans le cycle de réfrigération d'étage supérieur (20) et le fluide frigorigène circulant dans le cycle de réfrigération d'étage inférieur (10) ; et
    un dispositif de commande (30) configuré pour commuter entre un fonctionnement à deux étages dans lequel à la fois le cycle de réfrigération d'étage supérieur (20) et le cycle de réfrigération d'étage inférieur (10) fonctionnent, et un fonctionnement à un seul étage dans lequel le cycle de réfrigération d'étage supérieur (20) s'arrête et le cycle de réfrigération d'étage inférieur (10) fonctionne, de manière à sélectionner l'un du fonctionnement à deux étages et du fonctionnement à un étage ayant un COP plus élevé, dans lequel
    du fluide frigorigène de CO2 est utilisé dans le cycle de réfrigération d'étage inférieur (10), et caractérisé en ce que le système comprend en outre :
    un réservoir de détente (32) est disposé dans le cycle de réfrigération d'étage inférieur (10), ou
    un second détendeur (34) est disposé dans une partie amont d'un tuyau de liquide (16) entre le condenseur d'étage inférieur (12) et le détendeur d'étage inférieur (13) de telle sorte que le fluide frigorigène passant à travers le tuyau de liquide (16) est remplacé par un fluide frigorigène gaz-liquide à deux phases.
  2. Système de cycle de réfrigération à deux étages selon la revendication 1, dans lequel
    chacun du condenseur d'étage supérieur (22) et du radiateur auxiliaire (15) est un échangeur de chaleur qui transfère de la chaleur à l'air extérieur, et
    le dispositif de commande (30) passe du fonctionnement à deux étages au fonctionnement à un seul étage dans un cas où une température d'air extérieur dans le fonctionnement à deux étages est inférieure ou égale à une température d'air extérieur de seuil servant de valeur de seuil pour déterminer un niveau du COP.
  3. Système de cycle de réfrigération à deux étages selon la revendication 1, dans lequel
    le dispositif de commande (30) passe du fonctionnement à deux étages au fonctionnement à un seul étage dans un cas où un taux de compression du cycle de réfrigération d'étage inférieur (10) ou du cycle de réfrigération d'étage supérieur (20) dans le fonctionnement à deux étages est inférieur ou égal à un taux de compression de seuil servant de valeur de seuil pour déterminer un niveau du COP.
  4. Système de cycle de réfrigération à deux étages selon la revendication 1, dans lequel
    le dispositif de commande (30) passe du fonctionnement à deux étages au fonctionnement à un seul étage dans un cas où un taux de compression déterminé sur la base d'une pression d'évaporation du cycle de réfrigération d'étage inférieur (10) et d'une pression de condensation du cycle de réfrigération d'étage supérieur (20) dans le fonctionnement à deux étages est inférieur ou égal à un taux de compression de seuil servant de valeur de seuil pour déterminer un niveau du COP.
  5. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 4, dans lequel
    chacun du condenseur d'étage supérieur (22) et du radiateur auxiliaire (15) est un échangeur de chaleur qui transfère de la chaleur à l'air extérieur, et
    le dispositif de commande (30) passe du fonctionnement à un seul étage au fonctionnement à deux étages dans un cas où une température d'air extérieur dans le fonctionnement à un seul étage est supérieure à une température de seuil d'air extérieur servant de valeur de seuil pour déterminer un niveau du COP.
  6. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 4, dans lequel
    le dispositif de commande (30) passe du fonctionnement à un seul étage au fonctionnement à deux étages dans un cas où un taux de compression du cycle de réfrigération d'étage inférieur (10) dans le fonctionnement à un seul étage est supérieur à un taux de compression de seuil servant de valeur de seuil pour déterminer un niveau du COP.
  7. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 4, dans lequel
    le dispositif de commande (30) passe du fonctionnement à un seul étage au fonctionnement à deux étages dans un cas où une haute pression qui est une pression de condensation du cycle de réfrigération d'étage inférieur (10) dans le fonctionnement à un seul étage dépasse une haute pression qui est une pression de condensation du cycle de réfrigération d'étage inférieur (10) lorsque le fonctionnement à deux étages passe au fonctionnement à un seul étage.
  8. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 7, dans lequel
    le condenseur d'étage supérieur (22) et le radiateur auxiliaire (15) sont intégrés et forment un radiateur intégré (42), et
    le système de cycle de réfrigération à deux étages comprend en outre un dispositif d'envoi d'air (43) configuré pour envoyer de l'air au radiateur intégré (42).
  9. Système de cycle de réfrigération à deux étages selon la revendication 8, dans lequel
    le radiateur intégré (42) comprend une pluralité d'ailettes de transfert de chaleur (40) espacées les unes des autres de manière à permettre à de l'air de passer entre elles et une pluralité de tubes de transfert de chaleur (41) agencés de manière à s'étendre à travers la pluralité d'ailettes de transfert de chaleur (40),
    la pluralité de tubes de transfert de chaleur (41) sont agencés dans une pluralité de colonnes s'alignant perpendiculairement à une direction de passage d'air et une pluralité de rangées s'alignant dans la direction de passage d'air, et
    parmi la pluralité de tubes de transfert de chaleur (41), une pluralité de tubes de transfert de chaleur (41) constituant le radiateur auxiliaire (15) sont intégrés dans une rangée.
  10. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 9, dans lequel le condenseur d'étage supérieur (22) et le radiateur auxiliaire (15) ont une zone de transfert de chaleur sensiblement identique.
  11. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 10, dans lequel
    du fluide frigorigène présentant une efficacité supérieure à celle du fluide frigorigène utilisé dans le cycle de réfrigération d'étage inférieur (10) est utilisé dans le cycle de réfrigération d'étage supérieur (20).
  12. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 11, dans lequel
    du fluide frigorigène de CO2 ou du fluide frigorigène mixte comprenant du CO2 est utilisé dans au moins l'un du cycle de réfrigération d'étage inférieur (10) et du cycle de réfrigération d'étage supérieur (20).
  13. Système de cycle de réfrigération à deux étages selon la revendication 12, dans lequel
    du fluide frigorigène choisi parmi R32, R410A, R134a, R404A, R407C, HFO1234yf, HFO1234ze, l'ammoniac, le propane et l'isobutane, ou un fluide frigorigène mixte est utilisé dans au moins l'un du cycle de réfrigération d'étage inférieur (10) et du cycle de réfrigération d'étage supérieur (20).
  14. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 13, dans lequel
    du fluide frigorigène de CO2 est utilisé dans le cycle de réfrigération d'étage inférieur (10), et
    le dispositif de commande (30) établit une haute pression, qui est une pression de condensation du cycle de réfrigération d'étage inférieur (10) dans le fonctionnement à un seul étage, à une pression critique ou moins.
  15. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 14, dans lequel
    du fluide frigorigène de CO2 est utilisé dans le cycle de réfrigération d'étage inférieur (10),
    chacun du condenseur d'étage supérieur (22) et du radiateur auxiliaire (15) est un échangeur de chaleur qui transfère de la chaleur à l'air extérieur, et
    le dispositif de commande (30) passe du fonctionnement à deux étages au fonctionnement à un seul étage dans un cas où une température de l'air extérieur dans le fonctionnement à deux étages est inférieure ou égale à 21 degrés C.
  16. Système de cycle de réfrigération à deux étages selon l'une quelconque des revendications 1 à 15, comprenant à la fois le réservoir de détente (32) et le second détendeur (34).
EP13883949.3A 2013-05-08 2013-05-08 Dispositif de réfrigération binaire Active EP2995885B1 (fr)

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
PCT/JP2013/062931 WO2014181399A1 (fr) 2013-05-08 2013-05-08 Dispositif de réfrigération binaire

Publications (3)

Publication Number Publication Date
EP2995885A1 EP2995885A1 (fr) 2016-03-16
EP2995885A4 EP2995885A4 (fr) 2017-02-01
EP2995885B1 true EP2995885B1 (fr) 2020-04-15

Family

ID=50768279

Family Applications (1)

Application Number Title Priority Date Filing Date
EP13883949.3A Active EP2995885B1 (fr) 2013-05-08 2013-05-08 Dispositif de réfrigération binaire

Country Status (4)

Country Link
EP (1) EP2995885B1 (fr)
JP (1) JP6125000B2 (fr)
CN (1) CN203615641U (fr)
WO (1) WO2014181399A1 (fr)

Families Citing this family (21)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP6117078B2 (ja) * 2013-10-22 2017-04-19 大和ハウス工業株式会社 熱源システム
WO2016018692A1 (fr) * 2014-07-31 2016-02-04 Carrier Corporation Système de refroidissement
EP3312524B1 (fr) * 2015-06-18 2020-06-03 Mitsubishi Electric Corporation Dispositif à cycle de réfrigération
EP4086536A1 (fr) 2016-05-03 2022-11-09 Carrier Corporation Système de réfrigération à récupération de chaleur
EP3472541B1 (fr) 2016-06-17 2023-04-05 Carrier Corporation Sous-refroidisseur mécanique avec supplément batterie
DE102016213679A1 (de) * 2016-07-26 2018-02-01 Efficient Energy Gmbh Wärmepumpensystem mit eingangsseitig und ausgangsseitig gekoppelten Wärmepumpenanordnungen
DE102016213680A1 (de) * 2016-07-26 2018-02-01 Efficient Energy Gmbh Wärmepumpensystem mit CO2 als erstem Wärmepumpenmedium und Wasser als zweitem Wärmepumpenmedium
JP6712766B2 (ja) * 2017-02-14 2020-06-24 パナソニックＩｐマネジメント株式会社 二元冷凍システム
CN107062668B (zh) * 2017-03-09 2022-06-07 深圳市艾特网能技术有限公司 制冷循环系统及其制冷方法
KR102042353B1 (ko) * 2018-07-24 2019-11-07 한국해양대학교 산학협력단 냉동기
CN109579337B (zh) * 2018-11-26 2021-02-12 江苏天舒电器有限公司 一种复叠式热风控制系统与方法
CN109520167A (zh) * 2019-01-02 2019-03-26 天津城建大学 一种用于大型空调系统的多级复叠式制冷系统
JP6980945B2 (ja) 2019-02-25 2021-12-15 Ａｔｓジャパン株式会社 冷媒制御システム、及び冷却システム
EP4027074A4 (fr) * 2019-09-04 2022-09-07 Daikin Industries, Ltd. Unité de compresseur et dispositif de réfrigération
KR102352380B1 (ko) 2020-03-10 2022-01-17 에이티에스 저팬 가부시키가이샤 냉매 제어 시스템 및 냉각 시스템
US11421918B2 (en) 2020-07-10 2022-08-23 Energy Recovery, Inc. Refrigeration system with high speed rotary pressure exchanger
US11397030B2 (en) * 2020-07-10 2022-07-26 Energy Recovery, Inc. Low energy consumption refrigeration system with a rotary pressure exchanger replacing the bulk flow compressor and the high pressure expansion valve
KR102411307B1 (ko) * 2020-11-26 2022-06-21 김흥식 의약품 저온보관용 이원냉동장치
CN117120783A (zh) * 2021-04-21 2023-11-24 三菱电机株式会社 二元制冷循环装置
CN115289704B (zh) * 2022-06-23 2023-10-13 北京京仪自动化装备技术股份有限公司 温控装置和温控方法
CN115289705B (zh) * 2022-06-23 2024-03-15 北京京仪自动化装备技术股份有限公司 温控系统及温控方法

Family Cites Families (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS59219665A (ja) * 1983-05-27 1984-12-11 株式会社日立製作所 ヒ−トポンプ装置
JP2000274848A (ja) * 1999-03-23 2000-10-06 Daikin Ind Ltd 二元冷凍装置
JP3604973B2 (ja) * 1999-09-24 2004-12-22 三洋電機株式会社 カスケード式冷凍装置
JP2007218459A (ja) * 2006-02-15 2007-08-30 Matsushita Electric Ind Co Ltd 冷凍サイクル装置および保冷庫
JP5054180B2 (ja) * 2010-11-04 2012-10-24 サンデン株式会社 ヒートポンプ式暖房装置
EP2642220A4 (fr) * 2010-11-15 2017-04-19 Mitsubishi Electric Corporation Congélateur

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
None *

Also Published As

Publication number Publication date
JP6125000B2 (ja) 2017-05-10
CN203615641U (zh) 2014-05-28
EP2995885A1 (fr) 2016-03-16
EP2995885A4 (fr) 2017-02-01
JPWO2014181399A1 (ja) 2017-02-23
WO2014181399A1 (fr) 2014-11-13

Similar Documents

Publication Publication Date Title
EP2995885B1 (fr) Dispositif de réfrigération binaire
JP5318099B2 (ja) 冷凍サイクル装置、並びにその制御方法
EP2615392B1 (fr) Pompe de chaleur en cascade
US9523520B2 (en) Air-conditioning apparatus
US10208987B2 (en) Heat pump with an auxiliary heat exchanger for compressor discharge temperature control
US9599395B2 (en) Refrigerating apparatus
US9593872B2 (en) Heat pump
EP2808621B1 (fr) Dispositif de climatisation
EP3006866A1 (fr) Climatiseur
JP5332604B2 (ja) 冷暖同時運転型空気調和装置
WO2014128830A1 (fr) Dispositif de conditionnement d'air
EP2902726B1 (fr) Système pour alimentation en eau chaude et pour conditionnement d'air combinés
JP2007240025A (ja) 冷凍装置
EP3109566B1 (fr) Dispositif de climatisation
JP2009264606A (ja) 冷凍装置
KR101901540B1 (ko) 공기 조화 장치
JP2009264605A (ja) 冷凍装置
EP2837901A1 (fr) Système de refroidissement et son procédé de commande
EP3106779A1 (fr) Dispositif de réfrigération
WO2017175299A1 (fr) Dispositif à cycle frigorifique
JP6253370B2 (ja) 冷凍サイクル装置
EP3683520A1 (fr) Dispositif à cycle de réfrigération et dispositif de chauffage de liquide le comprenant
EP3165844B1 (fr) Dispositif de climatisation
JP2020046157A (ja) 冷凍装置
US20220412622A1 (en) Refrigeration cycle apparatus

Legal Events

Date Code Title Description
PUAI Public reference made under article 153(3) epc to a published international application that has entered the european phase

Free format text: ORIGINAL CODE: 0009012

17P Request for examination filed

Effective date: 20151120

AK Designated contracting states

Kind code of ref document: A1

Designated state(s): AL AT BE BG CH CY CZ DE DK EE ES FI FR GB GR HR HU IE IS IT LI LT LU LV MC MK MT NL NO PL PT RO RS SE SI SK SM TR

AX Request for extension of the european patent

Extension state: BA ME

DAX Request for extension of the european patent (deleted)
A4 Supplementary search report drawn up and despatched

Effective date: 20170105

RIC1 Information provided on ipc code assigned before grant

Ipc: F25B 7/00 20060101AFI20161223BHEP

Ipc: F25B 49/02 20060101ALI20161223BHEP

Ipc: F25B 6/04 20060101ALI20161223BHEP

Ipc: F25B 9/00 20060101ALI20161223BHEP

STAA Information on the status of an ep patent application or granted ep patent

Free format text: STATUS: EXAMINATION IS IN PROGRESS

17Q First examination report despatched

Effective date: 20190122

GRAP Despatch of communication of intention to grant a patent

Free format text: ORIGINAL CODE: EPIDOSNIGR1

STAA Information on the status of an ep patent application or granted ep patent

Free format text: STATUS: GRANT OF PATENT IS INTENDED

INTG Intention to grant announced

Effective date: 20191220

GRAS Grant fee paid

Free format text: ORIGINAL CODE: EPIDOSNIGR3

GRAA (expected) grant

Free format text: ORIGINAL CODE: 0009210

STAA Information on the status of an ep patent application or granted ep patent

Free format text: STATUS: THE PATENT HAS BEEN GRANTED

AK Designated contracting states

Kind code of ref document: B1

Designated state(s): AL AT BE BG CH CY CZ DE DK EE ES FI FR GB GR HR HU IE IS IT LI LT LU LV MC MK MT NL NO PL PT RO RS SE SI SK SM TR

REG Reference to a national code

Ref country code: CH

Ref legal event code: EP

REG Reference to a national code

Ref country code: DE

Ref legal event code: R096

Ref document number: 602013068057

Country of ref document: DE

REG Reference to a national code

Ref country code: IE

Ref legal event code: FG4D

REG Reference to a national code

Ref country code: AT

Ref legal event code: REF

Ref document number: 1257782

Country of ref document: AT

Kind code of ref document: T

Effective date: 20200515

REG Reference to a national code

Ref country code: NL

Ref legal event code: MP

Effective date: 20200415

REG Reference to a national code

Ref country code: LT

Ref legal event code: MG4D

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: SE

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: IS

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200815

Ref country code: NO

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200715

Ref country code: GR

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200716

Ref country code: FI

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: LT

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: NL

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: PT

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200817

REG Reference to a national code

Ref country code: AT

Ref legal event code: MK05

Ref document number: 1257782

Country of ref document: AT

Kind code of ref document: T

Effective date: 20200415

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: LV

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: RS

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: HR

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: BG

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200715

REG Reference to a national code

Ref country code: DE

Ref legal event code: R119

Ref document number: 602013068057

Country of ref document: DE

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: AL

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: EE

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: LI

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20200531

Ref country code: SM

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: IT

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: DK

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: CH

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20200531

Ref country code: MC

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: ES

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: AT

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: CZ

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: RO

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

PLBE No opposition filed within time limit

Free format text: ORIGINAL CODE: 0009261

STAA Information on the status of an ep patent application or granted ep patent

Free format text: STATUS: NO OPPOSITION FILED WITHIN TIME LIMIT

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: PL

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: SK

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

REG Reference to a national code

Ref country code: BE

Ref legal event code: MM

Effective date: 20200531

26N No opposition filed

Effective date: 20210118

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: LU

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20200508

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: IE

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20200508

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: DE

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20201201

Ref country code: BE

Free format text: LAPSE BECAUSE OF NON-PAYMENT OF DUE FEES

Effective date: 20200531

Ref country code: SI

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: TR

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: MT

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

Ref country code: CY

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

PG25 Lapsed in a contracting state [announced via postgrant information from national office to epo]

Ref country code: MK

Free format text: LAPSE BECAUSE OF FAILURE TO SUBMIT A TRANSLATION OF THE DESCRIPTION OR TO PAY THE FEE WITHIN THE PRESCRIBED TIME-LIMIT

Effective date: 20200415

REG Reference to a national code

Ref country code: FR

Ref legal event code: PLFP

Year of fee payment: 11

PGFP Annual fee paid to national office [announced via postgrant information from national office to epo]

Ref country code: GB

Payment date: 20230330

Year of fee payment: 11

P01 Opt-out of the competence of the unified patent court (upc) registered

Effective date: 20230512

PGFP Annual fee paid to national office [announced via postgrant information from national office to epo]

Ref country code: FR

Payment date: 20230411

Year of fee payment: 11