EP2290304A1 - Klimaanlage - Google Patents

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Publication number
EP2290304A1
EP2290304A1 EP09729048A EP09729048A EP2290304A1 EP 2290304 A1 EP2290304 A1 EP 2290304A1 EP 09729048 A EP09729048 A EP 09729048A EP 09729048 A EP09729048 A EP 09729048A EP 2290304 A1 EP2290304 A1 EP 2290304A1
Authority
EP
European Patent Office
Prior art keywords
pressure
temperature
heat exchanger
expansion valve
refrigerant
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.)
Withdrawn
Application number
EP09729048A
Other languages
English (en)
French (fr)
Other versions
EP2290304A4 (de
Inventor
Keisuke Hokazono
Takeshi Hatomura
Hiroyuki Morimoto
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 EP2290304A1 publication Critical patent/EP2290304A1/de
Publication of EP2290304A4 publication Critical patent/EP2290304A4/de
Withdrawn legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • 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
    • 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
    • F25B13/00—Compression machines, plants or systems, with reversible cycle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F25B41/30—Expansion means; Dispositions thereof
    • F25B41/39—Dispositions with two or more expansion means arranged in series, i.e. multi-stage expansion, on a refrigerant line leading to the same evaporator
    • 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
    • F25B2309/00—Gas cycle refrigeration machines
    • F25B2309/06—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide
    • F25B2309/061—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide with cycle highest pressure above the supercritical pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2313/00—Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/006—Compression machines, plants or systems with reversible cycle not otherwise provided for two pipes connecting the outdoor side to the indoor side with multiple indoor units
    • 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
    • F25B2313/00—Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/023—Compression machines, plants or systems with reversible cycle not otherwise provided for using multiple indoor units
    • F25B2313/0231—Compression machines, plants or systems with reversible cycle not otherwise provided for using multiple indoor units with simultaneous cooling and heating
    • 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
    • F25B2313/00—Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/027—Compression machines, plants or systems with reversible cycle not otherwise provided for characterised by the reversing means
    • F25B2313/02741—Compression machines, plants or systems with reversible cycle not otherwise provided for characterised by the reversing means using one four-way valve
    • 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
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/13—Economisers
    • 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
    • F25B2600/00—Control issues
    • F25B2600/11—Fan speed control
    • 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
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2509—Economiser valves
    • 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
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2513—Expansion valves
    • 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
    • F25B2700/19—Pressures
    • F25B2700/191—Pressures near an expansion valve

Definitions

  • the present invention relates to an air conditioner in which an outdoor unit and a plurality of indoor units are connected through a branch controller, and a supercritical fluid is used, to thereby establish a single refrigerating cycle.
  • the branch kit for each indoor unit is incorporated into a single branch controller for the purpose of reduction in number of connection pipes.
  • the air conditioner using the supercritical fluid has such a characteristic that the highest performance is obtained with a lower flow rate of the fluid by lowering temperature of the fluid to be conveyed to an indoor unit in cooling operation and raising temperature of the fluid to be conveyed to an indoor unit in heating operation. Accordingly, the efficiency (in this case, coefficient of performance (COP) expressed by taking performance of an air handling unit (unit: kW) as its numerator and power consumption (unit: kW) as its denominator) is enhanced as well.
  • COP coefficient of performance
  • temperature of an inlet of the indoor unit that is, temperature of an outlet of a heat-source side heat exchanger is basically lowered at the time of cooling and raised at the time of heating.
  • the refrigerating cycle is a refrigerating cycle of simultaneous heating and cooling operation in a cooling cycle
  • both cooling and heating are inevitably controlled at a certain degree of temperature (for example, approximately 40°C to 50°C under a pressure of 10 MPa in the supercritical range of the Mollier chart) of the outlet of the heat-source side heat exchanger. Consequently, there is an insufficiency of an enthalpy difference to obtain high performance, and hence a flow rate of the fluid is increased (power consumption of a compressor is increased) for compensation therefor, which results in a lower COP.
  • the efficiency of the air conditioner is conventionally evaluated based on the above-mentioned coefficient called COP in terms of only the efficiency with respect to 100% loads.
  • COP coefficient of the efficiency with respect to 100% loads.
  • cooling loads have been generated even in a season that necessitates heating, along with development of OA appliances and improvement in heat insulation performance of buildings.
  • the frequency of the simultaneous heating and cooling operation is becoming higher throughout a year. Therefore, the efficiency improvement tends more increasingly to be evaluated based not only on the COP with respect to the 100% loads but also on a COP obtained at the time of the simultaneous heating and cooling operation.
  • the conventional air conditioner has the problem of decline in COP caused by the operation for satisfying both cooling and heating.
  • the present invention has been made in view of the points described above, and it is therefore an object of the present invention to provide an air conditioner capable of improving a COP in simultaneous heating and cooling operation.
  • An air conditioner is an air conditioning system in which an outdoor unit and a plurality of indoor units are connected through a branch controller, and a supercritical fluid is used, to thereby establish a single refrigerating cycle, the outdoor unit and the branch controller being connected through two pipes of a high-pressure pipe and a low-pressure pipe, the branch controller and each of the plurality of indoor units being connected through two pipes of a high-pressure pipe and a low-pressure pipe, in which the branch controller includes a double-pipe heat exchanger for heat exchange between a medium-pressure two-phase refrigerant and a low-pressure two-phase refrigerant, the medium-pressure two-phase refrigerant being relatively high in temperature and flowing into the double-pipe heat exchanger after branching a refrigerant flowing from the outdoor unit toward the plurality of indoor units, and joining together a refrigerant decompressed by a first expansion valve and a refrigerant flowing from the plurality of indoor units, the low-pressure two-phase refrigerant being relatively low
  • connection pipes significant reduction in number of connection pipes is realized between the outdoor unit and the branch controller, and between the branch controller and each of the indoor units.
  • a COP in simultaneous heating and cooling operation is improved because a large enthalpy difference is secured on the side of the indoor units in cooling operation.
  • FIG 1 A diagram of a refrigerant circuit in cooling-dominant operation of an air conditioner according to a first embodiment of the present invention.
  • FIG 1 is a diagram of a refrigerant circuit in cooling-dominant operation of an air conditioner according to a first embodiment of the present invention.
  • an outdoor unit 100 and a plurality of indoor units 301 to 303 are connected through a branch controller 200, and a single refrigerating cycle is established by using a supercritical fluid.
  • the outdoor unit 100 mainly includes a compressor 110, a four-way valve 120, a heat-source side heat exchanger 130, and check valves 141 to 147.
  • the indoor units 301 to 303 respectively include use-side (load-side) heat exchangers 311 to 313, and expansion valves 321 to 323 serving as restriction devices.
  • the branch controller 200 mainly includes a first expansion valve 211, a second expansion valve 212, check valves 231 to 233, channel switching valves 221 to 223, and a double-pipe heat exchanger 240.
  • the double-pipe heat exchanger 240 may be a plate heat exchanger or a microchannel heat exchanger.
  • two pipes of a high-pressure pipe 400 and a low-pressure pipe 500 connect the outdoor unit 100 and the branch controller 200
  • two pipes of a high-pressure pipe 700 and a low-pressure pipe 800 similarly connect the branch controller 200 and each of the indoor units 301 to 303.
  • Cooling-dominant operation mainly involving cooling operation and partially involving heating operation is herein described.
  • heating-dominant operation channels are switched by means of the four-way valve 120 and the check valves 141 to 147.
  • FIG. 1 illustrates a high-pressure detection means 281, a medium-pressure detection means 282, a first temperature detection means 291, and a second temperature detection means 292 that are included in the branch controller 200, those components are used in a second embodiment described later but not necessary in the first embodiment.
  • the fluid of the outlet of the heat-source side heat exchanger 130 enters into a state of high pressure and medium temperature.
  • the fluid that flows out of the heat-source side heat exchanger 130 then flows into the branch controller 200 through the high-pressure pipe 400, and is branched into the indoor units 302 and 303 in cooling operation and the indoor unit 301 in heating operation by the respective channel switching valves 221 to 223.
  • the high-pressure and medium-temperature fluid that flows into the load-side heat exchanger 311 from a branch port through the channel switching valve 223 is further subjected to heat exchange with room temperature to become a high-pressure and medium-temperature fluid, which has temperature substantially equal to the room temperature (point C of FIG. 2 ). Then, the fluid is decompressed by the expansion valve 321 (point D of FIG. 2 ).
  • the refrigerant that flows out of the indoor unit 301 in heating operation through the low-pressure pipe 800 passes through the check valve 231 of the branch controller 200 in the state of medium pressure and medium temperature, and joins another refrigerant at a point between the first expansion valve 211 and the double-pipe heat exchanger 240.
  • the refrigerant for the side of the indoor units 302 and 303 in cooling operation flows from the branch port and is decompressed by the first expansion valve 211 down to medium pressure in the supercritical range, which is slightly lower than high pressure (point E of FIG. 2 ).
  • the refrigerant then flows into a medium-temperature side of the double-pipe heat exchanger 240 in the state of medium pressure and medium temperature.
  • the medium-pressure and medium-temperature fluid that is decompressed by the expansion valve 321 of the indoor unit 301 in heating operation joins this fluid at this point, and then flows into the medium-temperature side of the double-pipe heat exchanger 240.
  • a part of the fluid that flows out of the medium-temperature side of the double-pipe heat exchanger 240 is further branched at a branch port, and is decompressed by the second expansion valve 212 to become a low-pressure and low-temperature fluid having two phases of gas and liquid (point I of FIG. 2 ). Then, the fluid flows into a low-temperature side of the double-pipe heat exchanger 240.
  • the low-pressure and low-temperature fluid on the low-temperature side enters into a state of low pressure and medium temperature, and a high dryness (point H of FIG. 2 ) after being subjected to heat exchange with the medium-pressure and medium-temperature fluid on the medium-temperature side in the double-pipe heat exchanger 240.
  • the medium-pressure and medium-temperature fluid on the medium-temperature side is further cooled to become a medium-pressure and medium-temperature fluid in a state of a low enthalpy (point D of FIG. 2 ).
  • the medium-pressure and medium-temperature fluid that is further cooled point D of FIG.
  • FIG. 3 is a diagram of a refrigerant circuit in heating-dominant operation of the air conditioner according to the first embodiment of the present invention.
  • the air conditioner illustrated in FIG. 3 has the same configuration as the configuration of the first embodiment illustrated in FIG. 1 .
  • a high-pressure and high-temperature fluid that is compressed by the compressor 110 flows into the branch controller 200 through the four-way valve 120, the check valve 145, and the high-pressure pipe 400. Then, the high-pressure and high-temperature fluid is branched into the indoor unit 303 in cooling operation and the indoor units 301 and 302 in heating operation at the respective channel switching valves 221 to 223 of the branch controller 200. Further, the first expansion valve 211 is fully closed to block the flow.
  • the refrigerant for the side of the indoor units 301 and 302 in heating operation flows into the load-side heat exchangers 311 and 312 from the branch port of the branch controller 200 through the channel switching valves 222 and 223 and the high-pressure pipes 700, and the high-pressure and medium-temperature fluid is further subjected to heat exchange with the room temperature to become a high-pressure and medium-temperature fluid, which has temperature substantially equal to the room temperature (point C of FIG. 2 ). Then, the fluid is decompressed by the expansion valves 321 and 322 to become a medium-pressure and medium-temperature fluid (point D of FIG. 2 ).
  • the refrigerant that is decompressed by the expansion valves 321 and 322 then flows into the branch controller 200 through the low-pressure pipes 800, and passes through the check valves 231 and 232 in the state of medium pressure and medium temperature to join another refrigerant at the point between the first expansion valve 211 and the double-pipe heat exchanger 240.
  • the refrigerant for the side of the indoor unit 303 in cooling operation flows into the load-side expansion valve 323 through the following path.
  • the fluid that flows into the medium-pressure and medium-temperature side of the double-pipe heat exchanger 240 from the indoor units 301 and 302 in heating operation through the low-pressure pipes 800 and the check valves 231 and 232 is further branched at the branch port, and a part of the fluid is further decompressed by the second expansion valve 212 to become a low-pressure and low-temperature fluid (point I of FIG. 2 ).
  • the fluid then flows into the low-temperature side of the double-pipe heat exchanger 240.
  • the fluid is subjected to heat exchange with the medium-pressure and medium-temperature fluid on the medium-temperature side by the double-pipe heat exchanger 240.
  • the low-pressure and low-temperature fluid on the low-temperature side enters into the state of low pressure and medium temperature, and a high dryness (point H of FIG. 2 ), while the medium-pressure and medium-temperature fluid on the medium-temperature side is further cooled to become a medium-pressure and medium-temperature fluid in a state of a low enthalpy (point D of FIG. 2 ).
  • the medium-pressure and medium-temperature fluid that is further cooled (point D of FIG. 2 ) is then further decompressed by the load-side expansion valve 323 to become a low-pressure and low-temperature fluid. Then, the fluid flows into the load-side heat exchanger 313 for heat exchange with the room temperature, to thereby enter into the state of low pressure and medium temperature, and a high dryness (point H of FIG. 2 ).
  • the low-pressure and medium-temperature fluid that flows out of the low-temperature side of the double-pipe heat exchanger 240 and the low-pressure and medium-temperature fluid that flows out of the load-side heat exchanger 313 join each other, and return to the outdoor unit 100 side through the low-pressure pipe 500, the heat-source side heat exchanger 130, and the four-way valve 120.
  • the single outdoor unit 100 and the single branch controller 200 are connected through two pipes, and the branch controller 200 and each of the plurality of indoor units 301 to 303 are connected through two pipes. Accordingly, significant reduction in number of connection pipes is realized between the branch controller 200 and each of the indoor units 301 to 303, and at the same time, the COP in simultaneous heating and cooling operation is improved because the large enthalpy difference is secured on the side of the indoor units 302 and 303 in cooling operation. In addition, power-save operation is realized also in the cooling-dominant operation mainly involving cooling and partially involving heating operation.
  • a configuration of the second embodiment is the same as the configurations of the first embodiment illustrated in FIGS. 1 and 3 . Further, in FIGS. 1 and 3 , the high-pressure detection means 281, the medium-pressure detection means 282, the first temperature detection means 291, and the second temperature detection means 292 are provided to the branch controller 200, which are unnecessary in the first embodiment.
  • the flow of the refrigerant of the second embodiment is the same as that of the first embodiment.
  • a control method for the first expansion valve 211 is described.
  • Table 1 shows overviews of control in each control mode (full cooling, cooling dominant, full heating, or heating dominant).
  • full cooling In the case of the full-cooling operation (hereinafter, abbreviated as "full cooling") in which all the indoor units 301 to 303 perform cooling operation, the first expansion valve 211 is fully opened and the flow rate is controlled based on loads only by the expansion valves 321 to 323 of the indoor units 301 to 303.
  • temperature sensors 311a, 312a, and 313a for detecting temperatures on the upper side of the load-side heat exchangers 311 to 313, respectively.
  • temperature sensors 311b, 312b, and 313b for detecting temperatures and pressure sensors 311c, 312c, and 313c for detecting pressures between the load-side heat exchangers 311 to 313 and the load-side expansion valves 321 to 323, respectively.
  • temperature differences between the temperature sensors 311a, 312a, and 313a, and the temperature sensors 311b, 312b, and 313b are calculated, respectively, and calculation results thereof are each set as a degree of superheat.
  • the degree of opening of each of the load-side expansion valves 321, 322, and 323 is adjusted so that the degree of superheat becomes a predetermined value, for example, approximately 2°C.
  • pressure difference control which is described later, is performed by using a pressure difference between the high-pressure detection means 281 and the medium-pressure detection means 282. Also in the cooling-dominant operation, when the load-side heat exchangers operate as evaporators, the above-mentioned degree of superheat is detected and the degree of opening of each of the load-side expansion valves is adjusted so that the degree of superheat becomes a predetermined value.
  • full heating in which all the indoor units 301 to 303 perform heating operation, the first expansion valve 211 is fully closed basically and the flow rate is controlled based on loads only by the expansion valves 321 to 323 of the indoor units.
  • the pressure difference control which is described later, is performed by using the pressure difference between the high-pressure detection means 281 and the medium-pressure detection means 282.
  • the temperature sensors 311b, 312b, and 313b for detecting temperatures and the pressure sensors 311c, 312c, and 313c for detecting pressures are provided between the load-side heat exchangers 311 to 313 and the load-side expansion valves 321 to 323, respectively.
  • a virtual condensation temperature T c is calculated by using Expression (1) based on the pressure values obtained by the pressure sensors 311c, 312c, and 313c.
  • a difference between a temperature T L obtained by each of the temperature sensors 311b, 312b, and 313b and the virtual condensation temperature T c (Tc-T L ) is obtained, and this value is set as a degree of subcooling SC.
  • the degree of opening of each of the load-side expansion valves 321, 322, and 323 is adjusted so that the degree of subcooling SC becomes a predetermined value, for example, approximately 5°C.
  • the first expansion valve 211 is fully closed basically, and the pressure difference control, which is described later, is performed by using the pressure difference between the high-pressure detection means 281 and the medium-pressure detection means 282.
  • start up of the compressor 110 triggers the first expansion valve 211 to start with an initial degree of opening L0 that is set in advance (Step S41).
  • the degree of opening of the first expansion valve 211 is controlled according to comparison between the pressure difference ⁇ P obtained based on detection values of the high-pressure detection means 281 and the medium-pressure detection means 282, and set values P1 and P2 (P1 ⁇ P2) that are set in advance.
  • the degree of opening of the first expansion valve 211 is increased by a predetermined degree of opening that is set in advance (Steps S43 ⁇ S44).
  • the degree of opening of the first expansion valve 211 is maintained at the current degree of opening (Steps S43 ⁇ S45 ⁇ S46).
  • the degree of opening of the first expansion valve 211 is decreased by a predetermined degree of opening that is set in advance (Steps S43 ⁇ S45 ⁇ S47 ⁇ S48).
  • the first expansion valve 211 is fully closed constantly and the flow rate is controlled based on loads only by the expansion valves 321 to 323 of the indoor units.
  • start up of the compressor 110 triggers the first expansion valve 211 to start with the fully closed state (Step S51).
  • the predetermined period U has elapsed from the start (Step S52)
  • the degree of opening of the first expansion valve 211 is controlled according to comparison between the pressure difference ⁇ P obtained based on detection values of the high-pressure detection means 281 and the medium-pressure detection means 282, and the set values P1 and P2 (P1 ⁇ P2) that are set in advance.
  • the degree of opening of the first expansion valve 211 is increased by a predetermined degree of opening that is set in advance (Steps S53 ⁇ S54).
  • the degree of opening of the first expansion valve 211 is maintained at the current degree of opening (Steps S53 ⁇ S55 ⁇ S56).
  • the degree of opening of the first expansion valve 211 is decreased by a predetermined degree of opening that is set in advance (Steps S53 ⁇ S55 ⁇ S57 ⁇ S58).
  • a configuration of a third embodiment is the same as the configurations of the first embodiment illustrated in FIGS. 1 and 3 . Further, in FIGS. 1 and 3 , the high-pressure detection means 281, the medium-pressure detection means 282, the first temperature detection means 291, and the second temperature detection means 292 are provided to the branch controller 200, which are unnecessary in the first embodiment.
  • the flow of the refrigerant of the third embodiment is the same as that of the first embodiment.
  • a control method for the second expansion valve 212 is described.
  • Table 1 shows overviews of control in each control mode (full cooling, cooling dominant, full heating, or heating dominant).
  • the second expansion valve 212 is fully opened and the flow rate is controlled based on loads by the expansion valves 321 to 323 of the indoor units. Then, the refrigerant subjected to heat exchange with the load side flows into the low-pressure line of the outdoor unit through the second expansion valve 212.
  • the second expansion valve 212 is fully closed basically, and the pressure difference control, which is described later, is performed by using the pressure difference between the high-pressure detection means 281 and the medium-pressure detection means 282.
  • Step S61 start up of the compressor 110, or the like triggers the second expansion valve 212 to start with the initial degree of opening L0 that is set in advance.
  • Step S62 a temperature difference ⁇ T (degree of superheat) is calculated based on detection values of the first temperature detection means 291 and the second temperature detection means 292, and the degree of opening of the second expansion valve 212 is controlled according to comparison between the temperature difference ⁇ T and values T1 and T2 (T1 ⁇ T2) that are set in advance.
  • the degree of opening of the second expansion valve 212 is increased by a predetermined degree of opening that is set in advance (Steps S63 ⁇ S64).
  • the degree of opening of the second expansion valve 212 is maintained at the current degree of opening (Steps S63 ⁇ S65 ⁇ S66).
  • the degree of opening of the second expansion valve 212 is decreased by a predetermined degree of opening that is set in advance (Steps S63 ⁇ S65 ⁇ S67 ⁇ S68).
  • the second expansion valve 212 is fully opened constantly and the flow rate is controlled based on loads by the expansion valves 321 to 323 of the indoor units. Then, the refrigerant subjected to heat exchange with the load side flows into the low-pressure line of the outdoor unit through the second expansion valve 212.
  • start up of the compressor 110 triggers the second expansion valve 212 to start with the fully closed state (Step S71).
  • the degree of opening of the second expansion valve 212 is controlled according to comparison between the pressure difference ⁇ P obtained based on detection values of the high-pressure detection means 281 and the medium-pressure detection means 282, and the set values P1 and P2 (P1 ⁇ P2) that are set in advance.
  • the degree of opening of the second expansion valve 212 is decreased by a predetermined degree of opening that is set in advance (Steps S73 ⁇ S74).
  • the degree of opening of the second expansion valve 212 is maintained at the current degree of opening (Steps S73 ⁇ S75 ⁇ S76).
  • the degree of opening of the second expansion valve 212 is increased by a predetermined degree of opening that is set in advance (Steps S73 ⁇ S75 ⁇ S77 ⁇ S78).
  • the heat-source side heat exchanger 130 operates as a condenser (radiator).
  • the cooling load exceeds the heating load, and hence a part of the heat radiation capability needs to be supplemented by the heat-source side heat exchanger 130. Therefore, the heat exchanger capacity needs to be increased and decreased by adjusting the fan speed and dividing the heat-source side heat exchanger 130.
  • the enthalpy is uniquely determined.
  • a represents an enthalpy H 1 of the inlet of the heat-source side heat exchanger 130
  • b an enthalpy H 2 of the outlet of the heat-source side heat exchanger 130 (inlet of the load-side heat exchanger in heating operation)
  • c an enthalpy H 3 of the inlet of the heat-source side heat exchanger 130.
  • the enthalpy H 3 of the outlet of the load-side heat exchanger (heating) may be determined by using the pressure sensor 311c. and the temperature sensor 311b provided to the outlet of the load-side heat exchanger.
  • the enthalpy H 2 of the outlet of the heat-source side heat exchanger is determined.
  • the enthalpy of the outlet of the heat-source side heat exchanger may be obtained by using the pressure sensor 900 and the temperature sensor 901.
  • the enthalpy obtained from Expression (3) is set as a target enthalpy H 2m .
  • H 2s the enthalpy measured by the pressure sensor 900 and the temperature sensor 901 .
  • a control means increases and decreases the rotation of the heat-source side fan (blower) so that -epsH 2 ⁇ H s ⁇ epsH 2 (it should be noted that -epsH 2 and epsH 2 denote a lower limit value and an upper limit value in an error range, respectively) becomes a predetermined value. Control is performed so that, in a case where -epsH 2 > ⁇ H s , the number of rotations of the fan is increased, while in a case where ⁇ H s >epsH 2 , the number of rotations of the fan is decreased.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Chemical Kinetics & Catalysis (AREA)
  • Air Conditioning Control Device (AREA)
  • Compression-Type Refrigeration Machines With Reversible Cycles (AREA)
  • Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
EP09729048.0A 2008-03-31 2009-03-31 Klimaanlage Withdrawn EP2290304A4 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
PCT/JP2008/056370 WO2009122512A1 (ja) 2008-03-31 2008-03-31 空気調和装置
PCT/JP2009/056655 WO2009123190A1 (ja) 2008-03-31 2009-03-31 空気調和装置

Publications (2)

Publication Number Publication Date
EP2290304A1 true EP2290304A1 (de) 2011-03-02
EP2290304A4 EP2290304A4 (de) 2013-06-05

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EP09729048.0A Withdrawn EP2290304A4 (de) 2008-03-31 2009-03-31 Klimaanlage

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EP (1) EP2290304A4 (de)
JP (1) JPWO2009123190A1 (de)
WO (2) WO2009122512A1 (de)

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WO2012120097A3 (en) * 2011-03-08 2013-03-21 Greenfield Master Ipco Limited Thermal energy system and method of operation
EP2833086A4 (de) * 2012-03-27 2015-12-02 Mitsubishi Electric Corp Klimaanlage
US9360236B2 (en) 2008-06-16 2016-06-07 Greenfield Master Ipco Limited Thermal energy system and method of operation
WO2016173497A1 (zh) * 2015-04-28 2016-11-03 广东美的暖通设备有限公司 多联机系统及其中压控制方法
US9556856B2 (en) 2007-07-06 2017-01-31 Greenfield Master Ipco Limited Geothermal energy system and method of operation
US9915247B2 (en) 2007-07-06 2018-03-13 Erda Master Ipco Limited Geothermal energy system and method of operation
CN109445494A (zh) * 2018-10-10 2019-03-08 珠海格力电器股份有限公司 一种焓差实验室干球工况自动控制方法
EP4379288A1 (de) * 2022-11-29 2024-06-05 Panasonic Intellectual Property Management Co., Ltd. Dampfkompressionskühlzyklusvorrichtung

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EP2905552B1 (de) * 2012-10-01 2019-04-17 Mitsubishi Electric Corporation Klimaanlagenvorrichtung
JP6413692B2 (ja) * 2014-11-21 2018-10-31 株式会社富士通ゼネラル 空気調和装置
WO2017179166A1 (ja) * 2016-04-14 2017-10-19 三菱電機株式会社 空気調和装置

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JP4288979B2 (ja) * 2003-03-27 2009-07-01 三菱電機株式会社 空気調和装置、及び空気調和装置の運転制御方法
JP2005345069A (ja) * 2004-06-07 2005-12-15 Mitsubishi Heavy Ind Ltd 空気調和装置及びその運転制御方法
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JP2006283989A (ja) * 2005-03-31 2006-10-19 Sanyo Electric Co Ltd 冷暖房システム
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See also references of WO2009123190A1 *

Cited By (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US9556856B2 (en) 2007-07-06 2017-01-31 Greenfield Master Ipco Limited Geothermal energy system and method of operation
US9915247B2 (en) 2007-07-06 2018-03-13 Erda Master Ipco Limited Geothermal energy system and method of operation
US9360236B2 (en) 2008-06-16 2016-06-07 Greenfield Master Ipco Limited Thermal energy system and method of operation
WO2012120097A3 (en) * 2011-03-08 2013-03-21 Greenfield Master Ipco Limited Thermal energy system and method of operation
JP2014510895A (ja) * 2011-03-08 2014-05-01 グリーンフィールド マスター アイピーシーオー リミテッド 熱エネルギーシステム及びその動作方法
US10309693B2 (en) 2011-03-08 2019-06-04 Erda Master Ipco Limited Thermal energy system and method of operation
US10921030B2 (en) 2011-03-08 2021-02-16 Erda Master Ipco Limited Thermal energy system and method of operation
EP2833086A4 (de) * 2012-03-27 2015-12-02 Mitsubishi Electric Corp Klimaanlage
WO2016173497A1 (zh) * 2015-04-28 2016-11-03 广东美的暖通设备有限公司 多联机系统及其中压控制方法
CN109445494A (zh) * 2018-10-10 2019-03-08 珠海格力电器股份有限公司 一种焓差实验室干球工况自动控制方法
EP4379288A1 (de) * 2022-11-29 2024-06-05 Panasonic Intellectual Property Management Co., Ltd. Dampfkompressionskühlzyklusvorrichtung

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WO2009123190A1 (ja) 2009-10-08
EP2290304A4 (de) 2013-06-05
WO2009122512A1 (ja) 2009-10-08
JPWO2009123190A1 (ja) 2011-07-28

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