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WO2009123190A1 - Air conditioner - Google Patents

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Publication number
WO2009123190A1
WO2009123190A1 PCT/JP2009/056655 JP2009056655W WO2009123190A1 WO 2009123190 A1 WO2009123190 A1 WO 2009123190A1 JP 2009056655 W JP2009056655 W JP 2009056655W WO 2009123190 A1 WO2009123190 A1 WO 2009123190A1
Authority
WO
WIPO (PCT)
Prior art keywords
pressure
temperature
expansion valve
refrigerant
heat exchanger
Prior art date
Application number
PCT/JP2009/056655
Other languages
French (fr)
Japanese (ja)
Inventor
外囿 圭介
傑 鳩村
裕之 森本
Original Assignee
三菱電機株式会社
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 三菱電機株式会社 filed Critical 三菱電機株式会社
Priority to JP2010505935A priority Critical patent/JPWO2009123190A1/en
Priority to EP09729048.0A priority patent/EP2290304A4/en
Publication of WO2009123190A1 publication Critical patent/WO2009123190A1/en

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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B9/00Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008Compression 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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B13/00Compression machines, plants or systems, with reversible cycle
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00Fluid-circulation arrangements
    • F25B41/30Expansion means; Dispositions thereof
    • F25B41/39Dispositions with two or more expansion means arranged in series, i.e. multi-stage expansion, on a refrigerant line leading to the same evaporator
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2309/00Gas cycle refrigeration machines
    • F25B2309/06Compression machines, plants or systems characterised by the refrigerant being carbon dioxide
    • F25B2309/061Compression machines, plants or systems characterised by the refrigerant being carbon dioxide with cycle highest pressure above the supercritical pressure
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2313/00Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/006Compression 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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2313/00Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/023Compression machines, plants or systems with reversible cycle not otherwise provided for using multiple indoor units
    • F25B2313/0231Compression machines, plants or systems with reversible cycle not otherwise provided for using multiple indoor units with simultaneous cooling and heating
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2313/00Compression machines, plants or systems with reversible cycle not otherwise provided for
    • F25B2313/027Compression machines, plants or systems with reversible cycle not otherwise provided for characterised by the reversing means
    • F25B2313/02741Compression machines, plants or systems with reversible cycle not otherwise provided for characterised by the reversing means using one four-way valve
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00General 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
    • F25B2400/13Economisers
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00Control issues
    • F25B2600/11Fan speed control
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00Control issues
    • F25B2600/25Control of valves
    • F25B2600/2509Economiser valves
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00Control issues
    • F25B2600/25Control of valves
    • F25B2600/2513Expansion valves
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19Pressures
    • F25B2700/191Pressures 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 by a shunt controller, and one refrigeration cycle is configured using a supercritical fluid.
  • a heat recovery type air conditioner that simultaneously cools and warms using a supercritical fluid such as CO 2 is known.
  • the outdoor unit and the branch kit are mainly connected by three pipes of a high pressure pipe, a low pressure pipe and a high temperature gas pipe.
  • the branch kit from the branch kit to the indoor unit is a two-pipe type.
  • connection pipes it is conceivable to reduce the number of connection pipes by incorporating a branch kit for each indoor unit in one shunt controller in order to reduce the connection pipes.
  • the air conditioner using a supercritical fluid it is most effective to lower the temperature of the fluid sent to the cooling operation indoor unit and raise the temperature of the fluid sent to the heating operation indoor unit. Realized with low fluid flow rate. For this reason, efficiency (here, COP: Coefficient of Performance) in which the numerator is the capacity of the air conditioner (unit: kW) and the denominator is power consumption (unit: kW) is improved. Therefore, the inlet temperature of the indoor unit, that is, the outlet temperature of the heat source side heat exchanger is basically low during cooling and high during heating.
  • -It is necessary to lower the outlet temperature of the heat source side heat exchanger in order to supply a low temperature fluid to the cooling operation indoor unit.
  • -It is necessary to increase the outlet temperature of the heat source side heat exchanger in order to supply a high-temperature fluid to the heating operation indoor unit.
  • the conventional cooling main operation (refrigeration cycle is simultaneous cooling and heating operation in the cooling cycle) is a heat source side heat exchanger with a certain degree of cooling and heating (for example, a pressure of 10 MPa in the supercritical region, around 40 to 50 ° C. in the Mollier diagram).
  • a certain degree of cooling and heating for example, a pressure of 10 MPa in the supercritical region, around 40 to 50 ° C. in the Mollier diagram.
  • the conventional air conditioner has a problem that COP is lowered when it is operated so as to satisfy both the cooling and heating conditions.
  • the present invention has been made in view of the above points, and an object thereof is to obtain an air conditioner that can improve COP in simultaneous cooling and heating operations.
  • the air conditioning apparatus is an air conditioning system in which an outdoor unit and a plurality of indoor units are connected by a shunt controller, and a single refrigeration cycle is configured using a supercritical fluid.
  • the outdoor unit, the shunt controller Are connected by two pipes of a high pressure pipe and a low pressure pipe, and are connected between the branch flow controller and the plurality of indoor units by two pipes of a high pressure pipe and a low pressure pipe, and the branch flow controller is connected to the outdoor unit.
  • the refrigerant from the indoor unit to the indoor unit is branched and the refrigerant decompressed by the first expansion valve and the refrigerant from the indoor unit merge and flow into the indoor unit.
  • a double pipe heat exchanger for exchanging heat with a relatively low-temperature, low-pressure two-phase refrigerant that is branched and decompressed by a second expansion valve to flow out to the outdoor unit.
  • the number of connecting pipes between the outdoor unit and the branch flow controller and between the branch flow controller and each indoor unit can be greatly reduced, and a large enthalpy difference on the cooling operation indoor unit side can be secured.
  • COP in simultaneous cooling and heating is also improved.
  • FIG. FIG. 1 is a refrigerant circuit diagram at the time of cooling main operation of the air-conditioning apparatus according to Embodiment 1 of the present invention.
  • an outdoor unit 100 and a plurality of indoor units 301 to 303 are connected by a shunt controller 200 to constitute one refrigeration cycle using a supercritical fluid.
  • the outdoor unit 100 mainly includes a compressor 110, a four-way valve 120, a heat source unit side heat exchanger 130, and check valves 141 to 147.
  • the indoor units 301 to 303 include use side (load side) heat exchangers 311 to 313 and expansion valves 321 to 323 as expansion devices.
  • the flow dividing controller 200 mainly includes a first expansion valve 211, a second expansion valve 212, check valves 231 to 233, flow path switching valves 221 to 223, and a double pipe heat exchanger 240.
  • the double tube heat exchanger 240 may be a plate heat exchanger or a microchannel heat exchanger.
  • the outdoor unit 100 and the branch flow controller 200 are connected by two pipes, a high pressure pipe 400 and a low pressure pipe 500, and the high pressure pipe 700 and the low pressure pipe 200 are similarly connected between the branch flow controller 200 and each of the indoor units 301 to 303.
  • Two pipes 800 are connected to each other.
  • the cooling operation is mainly performed, and a cooling-main operation (hereinafter abbreviated as a cooling main operation) of a part of the heating operation is described, but the heating main operation (hereinafter abbreviated as a warm main operation) is described as a four-way valve 120,
  • the flow path is switched by check valves 141 to 147.
  • a high pressure detection means 281, an intermediate pressure detection means 282, a first temperature detection means 291, and a second temperature detection means 292 are shown in the shunt controller 200. Is unnecessary, and is used in Embodiment 2 to be described later.
  • the high-pressure and high-temperature fluid compressed by the compressor 110 is heat-exchanged with the surrounding air through the four-way valve 120 in the heat source unit-side heat exchanger 130 and cooled to a temperature that does not reach the ambient air temperature.
  • the degree of dryness of the Mollier diagram (pressure p-enthalpy h) shown in FIG. 2 is cooled to a temperature in the vicinity of 0.5 (point B in FIG. 2), and the outlet of the heat source side heat exchanger 130 is in a high pressure / intermediate temperature state. .
  • the fluid that has exited the heat source unit side heat exchanger 130 flows into the diversion controller 200 via the high-pressure pipe 400, and the cooling operation indoor units 302 and 303 and the heating operation chamber are flown at the flow path switching valves 221 to 223, respectively. Branch to machine 301.
  • the high-pressure / medium-temperature fluid that has flowed into the load-side heat exchanger 311 from the branch port via the flow path switching valve 223 further exchanges heat with room temperature.
  • the temperature becomes medium (point C in FIG. 2), and the pressure is reduced by the expansion valve 321 (point D in FIG. 2).
  • the refrigerant that has exited the heating operation indoor unit 301 via the low-pressure pipe 800 passes between the first expansion valve 211 and the double-pipe heat exchanger 240 via the check valve 231 in the shunt controller 200 in a state of intermediate pressure and intermediate temperature. Join at.
  • the refrigerant toward the cooling operation indoor units 302 and 303 is reduced from the branch port to the intermediate pressure in the supercritical region slightly lower than the high pressure by the first expansion valve 211 (point E in FIG. 2). It flows into the middle temperature side in the double-pipe heat exchanger 240 in the middle temperature state. Furthermore, the medium-pressure / medium-temperature fluid decompressed by the expansion valve 321 of the heating operation indoor unit 301 joins here and flows into the middle temperature side of the double-pipe heat exchanger 240.
  • the partial fluid that has come out from the middle temperature side of the double-pipe heat exchanger 240 is further branched at the branch port, and is further depressurized by the second expansion valve 212 to become a gas-liquid two-phase low-pressure low-temperature (I in FIG. 2). Point) and flows into the low temperature side in the double-tube heat exchanger 240.
  • the low-pressure low-temperature fluid at the low-temperature side becomes a state of low pressure and medium-temperature dryness (point H in FIG. 2).
  • the medium-pressure medium-temperature fluid on the medium-temperature side is further cooled to become a medium-pressure medium-temperature fluid (point D in FIG. 2) in a low enthalpy state.
  • the further cooled medium pressure medium temperature fluid (point D in FIG. 2) is further depressurized by the expansion valves 322 and 323 on the load side to become a gas-liquid two-phase low pressure and low temperature (point G in FIG. 2).
  • the number of connecting pipes between the outdoor unit 100 and the shunt controller 200 and between the shunt controller 200 and each of the indoor units 301 to 303 can be greatly reduced, and the cooling operation indoor units 302 and 303 side can be reduced.
  • COP during simultaneous cooling and heating is improved.
  • FIG. 3 is a refrigerant circuit diagram at the time of heating main operation of the air-conditioning apparatus according to Embodiment 1 of the present invention.
  • the air conditioning apparatus shown in FIG. 3 has the same configuration as that of the first embodiment shown in FIG.
  • FIG. 3 explains the flow in the refrigerant circuit during the warm main operation.
  • the high-pressure and high-temperature fluid compressed by the compressor 110 flows into the shunt controller 200 via the four-way valve 120, the check valve 145 and the high-pressure pipe 400. Further, the high-pressure and high-temperature fluid branches to the cooling operation indoor units 303 and the heating operation indoor units 301 and 302 at the flow path switching valves 221 to 223 in the flow dividing controller 200, respectively. Further, the flow of the first expansion valve 211 is interrupted in a fully closed state.
  • Refrigerant to the heating operation indoor units 301 and 302 side flows into the load-side heat exchangers 311 and 312 from the branch port in the diversion controller 200 via the flow path switching valves 222 and 223 and the high-pressure pipe 700, and the high-pressure intermediate temperature.
  • the fluid further exchanges heat with room temperature, and becomes a high pressure / intermediate temperature substantially equal to the room temperature (point C in FIG. 2).
  • the refrigerant depressurized by the expansion valves 321 and 322 flows into the diversion controller 200 via the low-pressure pipe 800 and exchanges heat with the first expansion valve 211 and the double pipe via the check valves 231 and 232.
  • the medium 240 is joined at a medium pressure and intermediate temperature state.
  • the refrigerant to the cooling operation indoor unit 303 side flows into the load side expansion valve 323 through the following path.
  • the fluid that has flowed from the heating operation indoor units 301 and 302 to the medium pressure / medium temperature side of the double-pipe heat exchanger 240 through the low-pressure pipe 800 and the check valves 231 and 232 is further branched at the branch port, and a part of the flow rate Is further reduced in pressure by the second expansion valve 212 to become a low pressure and low temperature (point I in FIG. 2), and flows into the low temperature side in the double pipe heat exchanger 240.
  • the low-temperature low-pressure low-temperature fluid is in a low-pressure / medium-temperature dryness state (point H in FIG.
  • the further cooled medium-pressure medium-temperature fluid (point D in FIG. 2) is further depressurized by the load-side expansion valve 323 to become low-pressure and low-temperature, and flows into the load-side heat exchanger 313, thereby As a result, heat is exchanged and the dryness of the low pressure medium temperature is large (point H in FIG. 2).
  • the low pressure / medium temperature fluid exiting the low temperature side of the double-pipe heat exchanger 240 and the low pressure / medium temperature fluid exiting the load side heat exchanger 313 merge to form a low pressure pipe 500, a heat source machine side heat exchanger. 130, Return to the outdoor unit 100 side through the four-way valve 120.
  • one outdoor unit 100 and one shunt controller 200 are connected by two pipes, and the shunt controller 200 and a plurality of indoor units 301 to 303 are connected by two pipes. Therefore, the number of connecting pipes from the shunt controller 200 to each of the indoor units 301 to 303 can be greatly reduced, and a large enthalpy difference on the cooling operation indoor units 302 and 303 side can be secured, thereby improving COP in simultaneous cooling and heating operations. To do.
  • an energy saving operation can be realized in a cooling main operation in which the cooling main operation is partly heating operation.
  • FIG. 2 the same configuration as that of the first embodiment shown in FIGS. 1 and 3 is provided. Further, in FIG. 1 and FIG. 3, the high pressure detection means 281, the intermediate pressure detection means 282, the first temperature detection means 291, and the second temperature detection means that are not required in the first embodiment are included in the shunt controller 200. 292 is provided.
  • Table 1 shows an outline of control in each control mode (full cooling, cooling main, total warming, warm main).
  • the first expansion valve 211 When all the indoor units 301 to 303 are in a cooling only operation (hereinafter, abbreviated as “fully cooled”), the first expansion valve 211 is fully opened, and the expansion valves of the indoor units 301 to 303 Only 321 to 323 performs flow control according to the load.
  • temperature sensors 311a, 312a, and 313a for detecting temperature are provided above the load side heat exchangers 311 to 313. Between the load-side heat exchangers 311 to 313 and the load-side expansion valves 321 to 323, temperature sensors 311b, 312b and 313b for detecting temperature and pressure sensors 311c, 312c and 313c for detecting pressure are provided. ing.
  • the temperature difference between the temperature sensors 311a, 312a, and 313a and the temperature sensors 311b, 312b, and 313b is calculated, and the calculation result is defined as a superheat degree (superheat), and the superheat degree is a predetermined value, for example,
  • the opening degree of the expansion valves 321, 322, and 323 on the load side is adjusted so as to be about 2 ° C.
  • differential pressure control which will be described later, using the differential pressures of the high pressure detection means 281 and the intermediate pressure detection means 282 is performed. Even in the cold main operation, when the load-side heat exchanger acts as an evaporator, the load-side expansion valve opening is adjusted to a predetermined degree of superheat while detecting the degree of superheat described above.
  • the first expansion valve 211 When all the indoor units 301 to 303 are in a heating operation that is a heating operation (hereinafter abbreviated as “fully warm”), the first expansion valve 211 is basically in a fully closed state, The flow rate control according to the load is performed only by the expansion valves 321 to 323 of the machine, and the differential pressure control described later using the differential pressure of the high pressure detecting means 281 and the intermediate pressure detecting means 282 is performed.
  • temperature sensors 311b, 312b and 313b for detecting temperature and pressure sensors 311c, 312c and 313c for detecting pressure are provided. Yes.
  • the saturation temperature T sc is calculated from the pressure values detected by the pressure sensors 311c, 312c, and 313c.
  • the calculated saturation temperature is defined as a condensation temperature Tc . It is necessary to prepare a relational expression between the saturation temperature T sc and the pressure P as shown in the formula (1) in advance.
  • T sc f (P) (1)
  • a virtual condensing temperature T c is calculated from the pressure values obtained from the pressure sensors 311b, 312b, and 313b using the equation (1).
  • a difference (T c ⁇ T L ) between the temperature T L obtained from the temperature sensors 311b, 312b, and 313b and the virtual condensing temperature T c is obtained, and this value is set as the degree of supercooling SC (subcool).
  • the opening degree of the expansion valves 321, 322, and 323 on the load side is adjusted so that the degree of supercooling SC becomes a predetermined value, for example, about 5 ° C.
  • the fully closed state is basically set, and the differential pressure between the high pressure detection means 281 and the intermediate pressure detection means 282 is changed.
  • the differential pressure control used later is performed.
  • the first expansion valve 211 is always in a fully opened state, and the flow rate control corresponding to the load is performed only by the expansion valves 321 to 323 of the indoor units.
  • the operation is started from a preset initial opening degree L0 (step S41), and a predetermined time from the start. After a lapse of U (step S42), according to the comparison between the differential pressure ⁇ P of the detected value of the high pressure detecting means 281 and the intermediate pressure detecting means 282 and the preset set values P1, P2 (P1 ⁇ P2) The opening degree of the 1 expansion valve 211 is controlled.
  • the first expansion valve 211 when ⁇ P> P2, the first expansion valve 211 is increased by a predetermined opening degree set in advance (steps S43 ⁇ S44), and when P1 ⁇ ⁇ P ⁇ P2, the first expansion valve 211 is opened at the current opening degree. (Step S43 ⁇ S45 ⁇ S46), and when P1 ⁇ P, the first expansion valve 211 is lowered by a predetermined opening degree (steps S43 ⁇ S45 ⁇ S47 ⁇ S48).
  • the first expansion valve 211 In the “fully warm” operation, the first expansion valve 211 is always in a fully closed state, and the flow rate is controlled according to the load only with the expansion valves 321 to 323 of the indoor unit.
  • the “warm main” operation starts from the fully closed state of the first expansion valve 211 with the start of the compressor 110 as a trigger (step S51), and starts a predetermined time U from the start.
  • the first expansion is performed in accordance with the comparison between the differential pressure ⁇ P between the detection values of the high pressure detection means 281 and the intermediate pressure detection means 282 and preset values P1, P2 (P1 ⁇ P2).
  • the opening degree of the valve 211 is controlled.
  • the first expansion valve 211 is increased by a predetermined opening degree set in advance (steps S53 ⁇ S54), and when P1 ⁇ ⁇ P ⁇ P2, the first expansion valve 211 is currently opened. (Step S53 ⁇ S55 ⁇ S56), and if P1 ⁇ P, the first expansion valve 211 is lowered by a predetermined opening degree (steps S53 ⁇ S55 ⁇ S57 ⁇ S58).
  • the above control ensures the necessary differential pressure to flow the refrigerant flow according to the load on the heating operation indoor unit side, and applies the differential pressure more than necessary, thereby reducing the inlet pressure of the cooling operation indoor unit.
  • the necessary differential pressure sufficient to allow the refrigerant flow rate corresponding to the load to flow in the cooling operation indoor unit can be secured, and as a result, it is possible to suppress a COP decrease.
  • a change in pressure can be suppressed, and the refrigerant can be stably sent to the indoor unit, so that energy saving operation and comfort can be realized.
  • Embodiment 3 FIG. In the third embodiment, the same configuration as the configuration of the first embodiment shown in FIG. 1 and FIG. Further, in FIG. 1 and FIG. 3, the high pressure detection means 281, the intermediate pressure detection means 282, the first temperature detection means 291, and the second temperature detection means that are not required in the first embodiment are included in the shunt controller 200. 292 is provided.
  • the refrigerant flow is the same as in the first embodiment.
  • Table 1 shows an outline of control in each control mode (full cooling, cool main, full warm, warm main).
  • the second expansion valve 212 uses the first temperature detection means 291 and the second temperature detection means 292, which will be described later. Superheat) control is performed, and on the indoor units 301 to 303 side, the expansion valves 321 to 323 perform flow rate control according to the load.
  • the “cooling main” operation in which the cooling and heating operations are performed simultaneously in the cooling cycle state is the same as “full cooling”.
  • the second expansion valve 212 When all the indoor units 301 to 303 are in the “fully warm” operation, which is a heating operation, the second expansion valve 212 is kept fully open, and the flow rate corresponding to the load is set by the expansion valves 321 to 323 of the indoor units.
  • the refrigerant that has been controlled and exchanged heat with the load side flows into the outdoor unit low-pressure line via the second expansion valve 212.
  • the differential pressure between the high pressure detection means 281 and the intermediate pressure detection means 282 is used as the fully closed state. Perform differential pressure control.
  • the “all-cooling” operation starts from the initial opening L0 set in advance by using the start of the compressor 110 as a trigger (step S61), and starts for a predetermined time from the start.
  • the temperature difference ⁇ T (superheat) of the detected values is calculated using the first temperature detecting means 291 and the second temperature detecting means 292, and the temperature difference ⁇ T is set to T1, which is set in advance.
  • the opening degree of the second expansion valve 212 is controlled according to the comparison with T2 (T1 ⁇ T2).
  • the second expansion valve 212 when ⁇ T> T2, the second expansion valve 212 is increased by a predetermined opening degree set in advance (steps S63 ⁇ S64), and when T1 ⁇ ⁇ T ⁇ T2, the second expansion valve 212 is currently opened. (Step S63 ⁇ S65 ⁇ S66), and when T1 ⁇ T, the second expansion valve 212 is lowered by a predetermined opening degree (steps S63 ⁇ S65 ⁇ S67 ⁇ S68).
  • the refrigerant temperature at the cooling operation indoor unit side inlet is cooled, and a necessary enthalpy difference sufficient to satisfy the performance can be secured, so that it is possible to suppress the decrease in COP. Further, even in the cooling main operation in which the cooling main operation is a partial heating operation, a lower temperature refrigerant can be sent to the cooling operation indoor unit, and an energy saving operation can be realized.
  • the second expansion valve 212 In the “fully warm” operation, the second expansion valve 212 is always fully opened, the flow rate is controlled according to the load by the expansion valves 321 to 323 of the indoor units, and the refrigerant exchanging heat with the load side is It flows into the outdoor unit low pressure line through the second expansion valve 212.
  • the “warm main” operation starts from the fully closed state triggered by the start of the compressor 110 (step S71), and after a predetermined time U has elapsed from the start (step S72).
  • the opening degree of the second expansion valve 212 is determined in accordance with a comparison between the pressure difference ⁇ P detected by the high pressure detection means 281 and the intermediate pressure detection means 282 and preset values P1, P2 (P1 ⁇ P2). To control.
  • the second expansion valve 212 is lowered by a predetermined opening degree set in advance (steps S73 ⁇ S74), and if P1 ⁇ ⁇ P ⁇ P2, the second expansion valve 212 is currently opened. (Step S73 ⁇ S75 ⁇ S76), and when P1 ⁇ P, the second expansion valve 212 is increased by a predetermined opening degree (steps S73 ⁇ S75 ⁇ S77 ⁇ S78).
  • the necessary differential pressure that allows the refrigerant flow rate to flow according to the load on the heating operation indoor unit side is secured, and the differential pressure more than necessary is applied (the intermediate pressure approaches low pressure), thereby cooling the system.
  • the inlet pressure of the operating indoor unit approaches a low pressure, and the required differential pressure sufficient to flow the refrigerant flow rate corresponding to the load cannot be secured in the cooling operation indoor unit, and as a result, it is possible to suppress a COP decrease.
  • the heat source machine side heat exchanger 130 acts with a condenser (heat radiator).
  • the cooling load is higher than the heating load, it is necessary to radiate a part of the heat radiation capability with the heat source unit side heat exchanger 130. For that purpose, it is necessary to divide the fan speed and the heat source device side heat exchanger 130 to increase or decrease the heat exchanger capacity.
  • a pressure sensor 900 and a temperature sensor 901 are provided between the heat source device side heat exchanger 130 and the check valve 141, and a temperature sensor 902 is provided at the inlet of the heat source device side heat exchanger 130. . In supercritical, if temperature and pressure are determined, enthalpy is uniquely determined.
  • a is the enthalpy H 1 at the inlet of the heat source machine side heat exchanger 130
  • b is the enthalpy H 2 at the outlet of the heat source machine side heat exchanger 130 (inlet of the load side heat exchanger in the heating operation)
  • c Is the enthalpy H 3 at the inlet of the heat source machine side heat exchanger 130.
  • Equation (3) Equation (3)
  • the enthalpy H 2 at the outlet of the heat source device side heat exchanger is determined.
  • the enthalpy at the heat source side heat exchanger outlet can be obtained from the pressure sensor 900 and the temperature sensor 901.
  • the enthalpy obtained from the equation (3) is set as the target enthalpy H 2m .
  • the enthalpy measured by the pressure sensor 900 and the temperature sensor 901 is assumed to be H 2s .
  • -epsH 2 ⁇ H s ⁇ epsH 2 (where -epsH 2 and epsH 2 indicate the lower limit value and the upper limit value of the error range) to be a predetermined value by the control means by the heat source side fan (blower) Increase or decrease the rotation of.
  • -epsH 2 > ⁇ H s the fan speed is increased, and when ⁇ H s > epsH 2 , the fan speed is decreased.

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Abstract

An air conditioner, the COPs of which in both cooling and heating operation are improved. In an air conditioning system in which an outdoor unit (100) and indoor units (301 to 303) are interconnected by a flow dividing controller (200) and a single refrigerating cycle is constructed by using a supercritical fluid, the outdoor unit (100) and the flow dividing controller (200) are interconnected by two lines of piping which are high-pressure piping (400) and low-pressure piping (500), and the flow dividing controller (200) and the indoor units (301 to 303) are interconnected by two lines of piping which are high-pressure piping (700) and low-pressure piping (800). The flow dividing controller (200) has a double tube heat exchanger (240) for exchanging heat between an intermediate pressure two-phase refrigerant and a low-pressure two-phase refrigerant. The intermediate pressure two-phase refrigerant is a refrigerant having a relatively high temperature, and refrigerants from the indoor units and a refrigerant which is generated by diverting a refrigerant from the outdoor unit to the indoor units and reducing this refrigerant in pressure by using a first expansion valve (211) are caused to merge into the intermediate pressure two-phase refrigerant. The low-pressure two-phase refrigerant is a refrigerant having a relatively low temperature and generated by diverting a refrigerant flowing into the indoor units, reducing the pressure of the refrigerant by a second expansion valve (212), and causing the refrigerant to flow out to the outdoor unit.

Description

空気調和装置Air conditioner
 この発明は、室外機と複数の室内機を分流コントローラにより接続し、超臨界流体を用いて1つの冷凍サイクルを構成した空気調和装置に関するものである。 The present invention relates to an air conditioner in which an outdoor unit and a plurality of indoor units are connected by a shunt controller, and one refrigeration cycle is configured using a supercritical fluid.
 従来、CO2などの超臨界流体を用いた冷暖同時の熱回収タイプの空気調和装置が知られている。このような空気調和装置においては、室外機と分岐キットを高圧配管と低圧配管及び高温ガス管の3管で接続されるのが主であった。なお、分岐キットから室内機までは2管式である。 2. Description of the Related Art Conventionally, a heat recovery type air conditioner that simultaneously cools and warms using a supercritical fluid such as CO 2 is known. In such an air conditioner, the outdoor unit and the branch kit are mainly connected by three pipes of a high pressure pipe, a low pressure pipe and a high temperature gas pipe. In addition, from the branch kit to the indoor unit is a two-pipe type.
 しかし、超臨界流体の臨界域における圧力は非常に高い圧力であるため、各ユニット間の接続配管の肉厚が従来のフロンに代表される冷媒の場合よりも大幅にアップする。このため、材料費コストの増大、または曲げなどの現地加工費が膨大に膨れ上がることが容易に予想される。 However, since the pressure in the critical region of the supercritical fluid is very high, the wall thickness of the connecting pipe between each unit is significantly increased compared to the case of refrigerants represented by conventional chlorofluorocarbons. For this reason, it is easily anticipated that the material cost increases or the local processing costs such as bending increase enormously.
 そこで、接続配管低減のために各室内機毎の分岐キットを1つの分流コントローラ内に内蔵し、接続配管本数を低減することが考えられる。 Therefore, it is conceivable to reduce the number of connection pipes by incorporating a branch kit for each indoor unit in one shunt controller in order to reduce the connection pipes.
 一方で、超臨界流体を用いた空気調和装置の特徴として、冷房運転室内機へ送り込む流体の温度を低く、暖房運転室内機へ送り込む流体の温度を高くすることが、最も能力が発揮され、より少ない流体の流量で実現される。このため、効率(ここでは、分子を空調機の能力(単位kW)、分母を消費電力(単位kW)としたCOP:Coefficient of Performance)もよくなる。従って、室内機の入口温度、すなわち、熱源側熱交換器の出口温度は、冷房時は低く、暖房時は高くするのが基本である。 On the other hand, as a feature of the air conditioner using a supercritical fluid, it is most effective to lower the temperature of the fluid sent to the cooling operation indoor unit and raise the temperature of the fluid sent to the heating operation indoor unit. Realized with low fluid flow rate. For this reason, efficiency (here, COP: Coefficient of Performance) in which the numerator is the capacity of the air conditioner (unit: kW) and the denominator is power consumption (unit: kW) is improved. Therefore, the inlet temperature of the indoor unit, that is, the outlet temperature of the heat source side heat exchanger is basically low during cooling and high during heating.
 しかし、2管式により冷暖同時を可能とする空気調和装置においては、冷房運転室内機と暖房運転室内機が同時に存在(混在)する場合において、以下のトレードオフが発生する。 However, in an air conditioner that enables simultaneous cooling and heating by a two-pipe system, the following trade-off occurs when the cooling operation indoor unit and the heating operation indoor unit exist (mixed) at the same time.
 ・冷房運転室内機に対しては、低い温度の流体を供給するために、熱源側熱交換器の出口温度を低くする必要がある。
 ・暖房運転室内機に対しては、高い温度の流体を供給するために、熱源側熱交換器の出口温度を高くする必要がある。
-It is necessary to lower the outlet temperature of the heat source side heat exchanger in order to supply a low temperature fluid to the cooling operation indoor unit.
-It is necessary to increase the outlet temperature of the heat source side heat exchanger in order to supply a high-temperature fluid to the heating operation indoor unit.
 例えば、従来での冷房主体運転(冷凍サイクルは冷房サイクルにおける冷暖同時運転)は、冷暖ともにある程度(例えば、モリエ線図で超臨界領域における圧力10MPa、40~50℃近辺)の熱源側熱交換器出口温度で制御せざるを得なく、結果的に能力を発揮するためには、エンタルピ差が不足し、その分流体流量を増やす(圧縮機消費電力アップ)ことで補い、その結果COPが低下する。 For example, the conventional cooling main operation (refrigeration cycle is simultaneous cooling and heating operation in the cooling cycle) is a heat source side heat exchanger with a certain degree of cooling and heating (for example, a pressure of 10 MPa in the supercritical region, around 40 to 50 ° C. in the Mollier diagram). In order to exert its ability as a result, it must be controlled by the outlet temperature, and the difference in enthalpy is insufficient, which is compensated by increasing the fluid flow rate (increasing the power consumption of the compressor), resulting in a decrease in COP. .
 さらに、空気調和装置の効率については、これまで前述のCOPと呼ばれる係数を用いて、100%負荷に対しての効率のみで評価を行っている。しかし、近年、例えば一般的な事務所における負荷は、OA機器の発達、および建築物の断熱性能の向上とともに、暖房シーズンにおいても冷房負荷が発生しており、年間を通じて冷暖同時運転の頻度が高まってきている。従って、100%負荷のCOPのみで評価するのではなく、冷暖同時運転でのCOPまで含めて効率改善する動向が強まってきている。 Furthermore, the efficiency of the air conditioner has been evaluated only with respect to 100% load by using the coefficient called COP described above. However, in recent years, for example, loads in general offices have been accompanied by the development of OA equipment and the improvement of the insulation performance of buildings, and cooling loads have also occurred in the heating season, and the frequency of simultaneous cooling and heating increases throughout the year. It is coming. Therefore, the trend of improving the efficiency including the COP in the simultaneous cooling and heating operation is increasing rather than evaluating only with the COP of 100% load.
 上述したように、従来の空気調和装置においては、冷暖房の両者を満たすように運転すると、COPが低下するという問題点があった。 As described above, the conventional air conditioner has a problem that COP is lowered when it is operated so as to satisfy both the cooling and heating conditions.
 この発明は上述した点に鑑みてなされたもので、冷暖同時運転でのCOPを改善できる空気調和装置を得ることを目的とする。 The present invention has been made in view of the above points, and an object thereof is to obtain an air conditioner that can improve COP in simultaneous cooling and heating operations.
 この発明に係る空気調和装置は、室外機と複数の室内機とを分流コントローラにより接続し、超臨界流体を用いて1つの冷凍サイクルを構成した空気調和システムにおいて、前記室外機と前記分流コントローラとの間を、高圧配管及び低圧配管の2管で接続し、前記分流コントローラと前記複数の室内機との間を、高圧配管及び低圧配管の2管で接続し、前記分流コントローラは、前記室外機から前記室内機への冷媒を分岐して第1膨張弁により減圧した冷媒と前記室内機からの冷媒とが合流して流入される比較的高温の中間圧二相冷媒と、前記室内機へ流出される冷媒を分岐して第2膨張弁により減圧して前記室外機へ流出される比較的低温の低圧二相冷媒とを熱交換させる二重管熱交換器を有することを特徴とする。 The air conditioning apparatus according to the present invention is an air conditioning system in which an outdoor unit and a plurality of indoor units are connected by a shunt controller, and a single refrigeration cycle is configured using a supercritical fluid. The outdoor unit, the shunt controller, Are connected by two pipes of a high pressure pipe and a low pressure pipe, and are connected between the branch flow controller and the plurality of indoor units by two pipes of a high pressure pipe and a low pressure pipe, and the branch flow controller is connected to the outdoor unit. The refrigerant from the indoor unit to the indoor unit is branched and the refrigerant decompressed by the first expansion valve and the refrigerant from the indoor unit merge and flow into the indoor unit. And a double pipe heat exchanger for exchanging heat with a relatively low-temperature, low-pressure two-phase refrigerant that is branched and decompressed by a second expansion valve to flow out to the outdoor unit.
 この発明によれば、室外機から分流コントローラとの間、および分流コントローラから各室内機との間までの接続配管本数を大幅に低減でき、かつ冷房運転室内機側でのエンタルピ差を大きく確保できることで冷暖同時運転でのCOPも向上する。 According to this invention, the number of connecting pipes between the outdoor unit and the branch flow controller and between the branch flow controller and each indoor unit can be greatly reduced, and a large enthalpy difference on the cooling operation indoor unit side can be secured. COP in simultaneous cooling and heating is also improved.
この発明の実施の形態1に係る空気調和装置の冷房主体運転時の冷媒回路図である。It is a refrigerant circuit figure at the time of the cooling main operation | movement of the air conditioning apparatus which concerns on Embodiment 1 of this invention. この発明の実施の形態1に係る空気調和装置の説明に用いるモリエ線図である。It is a Mollier diagram used for description of the air conditioning apparatus which concerns on Embodiment 1 of this invention. この発明の実施の形態1に係る空気調和装置の暖房主体運転時の冷媒回路図である。It is a refrigerant circuit figure at the time of heating main operation | movement of the air conditioning apparatus which concerns on Embodiment 1 of this invention. この発明の実施の形態2に係る空気調和装置における冷主運転時の第1膨張弁211の制御フローチャートである。It is a control flowchart of the 1st expansion valve 211 at the time of the cold main operation | movement in the air conditioning apparatus which concerns on Embodiment 2 of this invention. この発明の実施の形態2に係る空気調和装置における全暖及び暖主運転時の第1膨張弁211の制御フローチャートである。It is a control flowchart of the 1st expansion valve 211 at the time of the full warm and warm main operation | movement in the air conditioning apparatus which concerns on Embodiment 2 of this invention. この発明の実施の形態3に係る空気調和装置における全冷及び冷主運転時の第2膨張弁212の制御フローチャートである。It is a control flowchart of the 2nd expansion valve 212 at the time of the total cooling in the air conditioning apparatus which concerns on Embodiment 3 of this invention, and cold main operation. この発明の実施の形態3に係る空気調和装置における暖主運転時の第2膨張弁212の制御フローチャートである。It is a control flowchart of the 2nd expansion valve 212 at the time of the warm main operation | movement in the air conditioning apparatus which concerns on Embodiment 3 of this invention. この発明の実施の形態3に係る空気調和装置の説明に用いるモリエ線図である。It is a Mollier diagram used for description of the air conditioning apparatus which concerns on Embodiment 3 of this invention.
 実施の形態1.
 図1は、この発明の実施の形態1に係る空気調和装置の冷房主体運転時の冷媒回路図である。図1に示す空気調和装置は、室外機100と複数の室内機301~303とが分流コントローラ200により接続され、超臨界流体を用いて1つの冷凍サイクルを構成している。室外機100は、主に、圧縮機110、四方弁120、熱源機側熱交換器130及び逆止弁141~147を具備している。また、室内機301~303は、利用側(負荷側)熱交換器311~313と、絞り装置としての膨張弁321~323を具備している。さらに、分流コントローラ200は、主に、第1膨張弁211、第2膨張弁212、逆止弁231~233、流路切換弁221~223及び二重管熱交換器240を具備している。なお、二重管熱交換器240は、プレート熱交換器またはマイクロチャネル熱交換器であってもよい。
Embodiment 1 FIG.
FIG. 1 is a refrigerant circuit diagram at the time of cooling main operation of the air-conditioning apparatus according to Embodiment 1 of the present invention. In the air conditioner shown in FIG. 1, an outdoor unit 100 and a plurality of indoor units 301 to 303 are connected by a shunt controller 200 to constitute one refrigeration cycle using a supercritical fluid. The outdoor unit 100 mainly includes a compressor 110, a four-way valve 120, a heat source unit side heat exchanger 130, and check valves 141 to 147. The indoor units 301 to 303 include use side (load side) heat exchangers 311 to 313 and expansion valves 321 to 323 as expansion devices. Further, the flow dividing controller 200 mainly includes a first expansion valve 211, a second expansion valve 212, check valves 231 to 233, flow path switching valves 221 to 223, and a double pipe heat exchanger 240. The double tube heat exchanger 240 may be a plate heat exchanger or a microchannel heat exchanger.
 ここで、室外機100と分流コントローラ200との間は、高圧配管400と低圧配管500の2管で接続され、分流コントローラ200と各室内機301~303との間も同様に高圧配管700と低圧配管800の2管でそれぞれ接続されている。ここでは、冷房運転が主体で、一部暖房運転の冷房主体運転(以下、冷主運転と略す)について記載するが、暖房主体運転(以下、暖主運転と略す)については、四方弁120、逆止弁141~147にて流路が切り換えられる。 Here, the outdoor unit 100 and the branch flow controller 200 are connected by two pipes, a high pressure pipe 400 and a low pressure pipe 500, and the high pressure pipe 700 and the low pressure pipe 200 are similarly connected between the branch flow controller 200 and each of the indoor units 301 to 303. Two pipes 800 are connected to each other. Here, the cooling operation is mainly performed, and a cooling-main operation (hereinafter abbreviated as a cooling main operation) of a part of the heating operation is described, but the heating main operation (hereinafter abbreviated as a warm main operation) is described as a four-way valve 120, The flow path is switched by check valves 141 to 147.
 なお、図1では、分流コントローラ200内に、高圧圧力検知手段281、中間圧圧力検知手段282、第1温度検知手段291、第2温度検知手段292が図示されているが、この実施の形態1では不要であり、後述する実施の形態2において用いられる。 In FIG. 1, a high pressure detection means 281, an intermediate pressure detection means 282, a first temperature detection means 291, and a second temperature detection means 292 are shown in the shunt controller 200. Is unnecessary, and is used in Embodiment 2 to be described later.
 まず、図1にて冷主運転での冷媒回路内の流れを説明する。ここでは、超臨界流体としてCO2を用いた場合について述べる。圧縮機110で圧縮された高圧高温の流体は、四方弁120を介して熱源機側熱交換器130にて周囲の空気と熱交換され、周囲の空気温度まではいかない温度まで冷却され、例えば、図2に示すモリエ線図(圧力p-エンタルピh)の乾き度が0.5近辺(図2中B点)の温度まで冷却され、熱源機側熱交換器130出口が高圧中温の状態となる。熱源機側熱交換器130を出た流体は、高圧配管400を介して分流コントローラ200内に流入し、その流路切換弁221~223にて、それぞれ冷房運転室内機302,303、暖房運転室内機301へと分岐する。 First, the flow in the refrigerant circuit in the cold main operation will be described with reference to FIG. Here, the case where CO 2 is used as the supercritical fluid will be described. The high-pressure and high-temperature fluid compressed by the compressor 110 is heat-exchanged with the surrounding air through the four-way valve 120 in the heat source unit-side heat exchanger 130 and cooled to a temperature that does not reach the ambient air temperature. The degree of dryness of the Mollier diagram (pressure p-enthalpy h) shown in FIG. 2 is cooled to a temperature in the vicinity of 0.5 (point B in FIG. 2), and the outlet of the heat source side heat exchanger 130 is in a high pressure / intermediate temperature state. . The fluid that has exited the heat source unit side heat exchanger 130 flows into the diversion controller 200 via the high-pressure pipe 400, and the cooling operation indoor units 302 and 303 and the heating operation chamber are flown at the flow path switching valves 221 to 223, respectively. Branch to machine 301.
 暖房運転室内機301側への冷媒は、分岐口から流路切換弁223を介して負荷側熱交換器内311へ流入した高圧中温の流体がさらに室温と熱交換し、室温とほぼ同等の高圧中温となり(図2中C点)、膨張弁321にて減圧される(図2中D点)。低圧配管800を介して暖房運転室内機301を出た冷媒は、中圧中温の状態で分流コントローラ200内の逆止弁231を介して、第1膨張弁211と二重管熱交換器240間で合流する。 As for the refrigerant to the heating operation indoor unit 301 side, the high-pressure / medium-temperature fluid that has flowed into the load-side heat exchanger 311 from the branch port via the flow path switching valve 223 further exchanges heat with room temperature. The temperature becomes medium (point C in FIG. 2), and the pressure is reduced by the expansion valve 321 (point D in FIG. 2). The refrigerant that has exited the heating operation indoor unit 301 via the low-pressure pipe 800 passes between the first expansion valve 211 and the double-pipe heat exchanger 240 via the check valve 231 in the shunt controller 200 in a state of intermediate pressure and intermediate temperature. Join at.
 一方、冷房運転室内機302,303側への冷媒は、前記分岐口から第1膨張弁211で高圧よりも少し低い超臨界域での中間圧まで減圧され(図2中E点)、中圧中温の状態で二重管熱交換器240内の中温側へ流入する。さらに、暖房運転室内機301の膨張弁321で減圧された中圧中温の流体もここで合流して二重管熱交換器240の中温側へ流入する。ここで、二重管熱交換器240の中温側から出た一部流体は分岐口でさらに分岐され、第2膨張弁212にてさらに減圧され気液二相の低圧低温となり(図2中I点)、二重管熱交換器240内の低温側へ流入する。 On the other hand, the refrigerant toward the cooling operation indoor units 302 and 303 is reduced from the branch port to the intermediate pressure in the supercritical region slightly lower than the high pressure by the first expansion valve 211 (point E in FIG. 2). It flows into the middle temperature side in the double-pipe heat exchanger 240 in the middle temperature state. Furthermore, the medium-pressure / medium-temperature fluid decompressed by the expansion valve 321 of the heating operation indoor unit 301 joins here and flows into the middle temperature side of the double-pipe heat exchanger 240. Here, the partial fluid that has come out from the middle temperature side of the double-pipe heat exchanger 240 is further branched at the branch port, and is further depressurized by the second expansion valve 212 to become a gas-liquid two-phase low-pressure low-temperature (I in FIG. 2). Point) and flows into the low temperature side in the double-tube heat exchanger 240.
 二重管熱交換器240にて前記中温側の中圧中温の流体と熱交換することで、低温側の低圧低温流体は、低圧中温の乾き度の大きい状態(図2中H点)となり、一方、中温側の中圧中温流体は、さらに冷却され、低エンタルピ状態の中圧中温流体(図2中D点)となる。そして、このさらに冷却された中圧中温流体(図2中D点)が、負荷側の膨張弁322,323にてさらに減圧されて気液二相の低圧低温(図2中G点)となり、負荷側の熱交換器322,323内へ流入することで、室温と熱交換し、低圧中温の乾き度の大きい状態となる(図2中H点)。最後に、二重管熱交換器240の低温側を出た低圧中温の流体と、負荷側熱交換器322,323を出た低圧中温の流体とが合流し、低圧配管500を介して室外機100側へ戻る。 By exchanging heat with the medium-pressure medium-temperature fluid at the medium-temperature side in the double-tube heat exchanger 240, the low-pressure low-temperature fluid at the low-temperature side becomes a state of low pressure and medium-temperature dryness (point H in FIG. 2). On the other hand, the medium-pressure medium-temperature fluid on the medium-temperature side is further cooled to become a medium-pressure medium-temperature fluid (point D in FIG. 2) in a low enthalpy state. The further cooled medium pressure medium temperature fluid (point D in FIG. 2) is further depressurized by the expansion valves 322 and 323 on the load side to become a gas-liquid two-phase low pressure and low temperature (point G in FIG. 2). By flowing into the heat exchangers 322 and 323 on the load side, heat exchange with room temperature occurs, and the dryness at low pressure and intermediate temperature is high (point H in FIG. 2). Finally, the low pressure / medium temperature fluid exiting the low temperature side of the double pipe heat exchanger 240 and the low pressure / medium temperature fluid exiting the load side heat exchangers 322 and 323 merge, and the outdoor unit is connected via the low pressure pipe 500. Return to the 100 side.
 これにより、室外機100から分流コントローラ200との間、および分流コントローラ200から各室内機301~303との間までの接続配管本数を大幅に低減でき、かつ冷房運転室内機302,303側でのエンタルピ差を大きく確保できたことで冷暖同時運転でのCOPも向上する。 As a result, the number of connecting pipes between the outdoor unit 100 and the shunt controller 200 and between the shunt controller 200 and each of the indoor units 301 to 303 can be greatly reduced, and the cooling operation indoor units 302 and 303 side can be reduced. By ensuring a large enthalpy difference, COP during simultaneous cooling and heating is improved.
 次に、図3は、この発明の実施の形態1に係る空気調和装置の暖房主体運転時の冷媒回路図である。図3に示す空気調和装置は、図1に示す実施の形態1の構成と同様な構成を備える。 Next, FIG. 3 is a refrigerant circuit diagram at the time of heating main operation of the air-conditioning apparatus according to Embodiment 1 of the present invention. The air conditioning apparatus shown in FIG. 3 has the same configuration as that of the first embodiment shown in FIG.
 図3により、暖主運転時の冷媒回路内の流れを説明する。圧縮機110で圧縮された高圧高温の流体は、四方弁120、逆止弁145および高圧配管400を介して、分流コントローラ200へ流入する。さらに、高圧高温の流体は、分流コントローラ200内の流路切換弁221~223にて、それぞれ冷房運転室内機303、暖房運転室内機301,302へと分岐する。また、第1膨張弁211は、全閉の状態で流れが遮断されている。 FIG. 3 explains the flow in the refrigerant circuit during the warm main operation. The high-pressure and high-temperature fluid compressed by the compressor 110 flows into the shunt controller 200 via the four-way valve 120, the check valve 145 and the high-pressure pipe 400. Further, the high-pressure and high-temperature fluid branches to the cooling operation indoor units 303 and the heating operation indoor units 301 and 302 at the flow path switching valves 221 to 223 in the flow dividing controller 200, respectively. Further, the flow of the first expansion valve 211 is interrupted in a fully closed state.
 暖房運転室内機301,302側への冷媒は、分流コントローラ200内の分岐口から流路切換弁222、223および高圧配管700を介して負荷側熱交換器内311,312へ流入し、高圧中温の流体がさらに室温と熱交換し、室温とほぼ同等の高圧中温となり(図2中C点)、膨張弁321,322にて減圧され、中圧中温となる(図2中D点)。そして、膨張弁321,322にて減圧された冷媒は、低圧配管800を介して分流コントローラ200内へ流入し、逆止弁231,232を介して、第1膨張弁211と二重管熱交換器240の間へ中圧中温状態で合流する。 Refrigerant to the heating operation indoor units 301 and 302 side flows into the load- side heat exchangers 311 and 312 from the branch port in the diversion controller 200 via the flow path switching valves 222 and 223 and the high-pressure pipe 700, and the high-pressure intermediate temperature. The fluid further exchanges heat with room temperature, and becomes a high pressure / intermediate temperature substantially equal to the room temperature (point C in FIG. 2). Then, the refrigerant depressurized by the expansion valves 321 and 322 flows into the diversion controller 200 via the low-pressure pipe 800 and exchanges heat with the first expansion valve 211 and the double pipe via the check valves 231 and 232. The medium 240 is joined at a medium pressure and intermediate temperature state.
 一方、冷房運転室内機303側への冷媒は、以下の経路を介して負荷側膨張弁323へ流入する。暖房運転室内機301,302から低圧配管800および逆止弁231,232を介して二重管熱交換器240の中圧中温側へ流入した流体は、分岐口でさらに分岐され、一部の流量が第2膨張弁212にてさらに減圧され低圧低温となり(図2中I点)、二重管熱交換器240内の低温側へ流入する。そして、二重管熱交換器240にて前記中温側の中圧中温の流体と熱交換することで、低温側の低圧低温流体は、低圧中温の乾き度の大きい状態(図2中H点)となり、中温側の中圧中温流体は、さらに冷却され、低エンタルピ状態の中圧中温流体(図2中D点)となる。 On the other hand, the refrigerant to the cooling operation indoor unit 303 side flows into the load side expansion valve 323 through the following path. The fluid that has flowed from the heating operation indoor units 301 and 302 to the medium pressure / medium temperature side of the double-pipe heat exchanger 240 through the low-pressure pipe 800 and the check valves 231 and 232 is further branched at the branch port, and a part of the flow rate Is further reduced in pressure by the second expansion valve 212 to become a low pressure and low temperature (point I in FIG. 2), and flows into the low temperature side in the double pipe heat exchanger 240. Then, the low-temperature low-pressure low-temperature fluid is in a low-pressure / medium-temperature dryness state (point H in FIG. 2) by exchanging heat with the medium-temperature medium-pressure medium-temperature fluid in the double-tube heat exchanger 240. Thus, the medium-pressure medium-temperature fluid at the medium-temperature side is further cooled to become a medium-pressure medium-temperature fluid (point D in FIG. 2) in a low enthalpy state.
 そして、このさらに冷却された中圧中温流体(図2中D点)が、負荷側の膨張弁323にてさらに減圧されて低圧低温となり、負荷側熱交換器313内へ流入することで、室温と熱交換され、低圧中温の乾き度の大きい状態となる(図2中H点)。最後に、二重管熱交換器240の低温側を出た低圧中温の流体と、負荷側熱交換器313を出た低圧中温の流体とが合流し、低圧配管500、熱源機側熱交換器130、四方弁120を介して室外機100側へ戻る。 The further cooled medium-pressure medium-temperature fluid (point D in FIG. 2) is further depressurized by the load-side expansion valve 323 to become low-pressure and low-temperature, and flows into the load-side heat exchanger 313, thereby As a result, heat is exchanged and the dryness of the low pressure medium temperature is large (point H in FIG. 2). Finally, the low pressure / medium temperature fluid exiting the low temperature side of the double-pipe heat exchanger 240 and the low pressure / medium temperature fluid exiting the load side heat exchanger 313 merge to form a low pressure pipe 500, a heat source machine side heat exchanger. 130, Return to the outdoor unit 100 side through the four-way valve 120.
 したがって、実施の形態1によれば、1台の室外機100と1台の分流コントローラ200とを2管で接続し、分流コントローラ200と複数の室内機301~303とを2管で接続することで、分流コントローラ200から各室内機301~303までの接続配管本数を大幅に低減でき、かつ冷房運転室内機302,303側でのエンタルピ差を大きく確保できることで、冷暖同時運転でのCOPも向上する。また、冷房主体で一部暖房運転となる冷房主体運転においても省エネ運転を実現することができる。 Therefore, according to the first embodiment, one outdoor unit 100 and one shunt controller 200 are connected by two pipes, and the shunt controller 200 and a plurality of indoor units 301 to 303 are connected by two pipes. Therefore, the number of connecting pipes from the shunt controller 200 to each of the indoor units 301 to 303 can be greatly reduced, and a large enthalpy difference on the cooling operation indoor units 302 and 303 side can be secured, thereby improving COP in simultaneous cooling and heating operations. To do. In addition, an energy saving operation can be realized in a cooling main operation in which the cooling main operation is partly heating operation.
 実施の形態2.
 この実施の形態2では、図1及び図3に示す実施の形態1の構成と同様な構成を備える。さらに、図1及び図3において、分流コントローラ200内に、実施の形態1では不要であった、高圧圧力検知手段281、中間圧圧力検知手段282、第1温度検知手段291、第2温度検知手段292が備えられる。
Embodiment 2. FIG.
In the second embodiment, the same configuration as that of the first embodiment shown in FIGS. 1 and 3 is provided. Further, in FIG. 1 and FIG. 3, the high pressure detection means 281, the intermediate pressure detection means 282, the first temperature detection means 291, and the second temperature detection means that are not required in the first embodiment are included in the shunt controller 200. 292 is provided.
 この実施の形態2において、冷媒の流れは、実施の形態1と同じである。以下、第1膨張弁211の制御方法について説明する。まず、各制御モード(全冷、冷主、全暖、暖主)における制御概要は、表1の通りである。 In the second embodiment, the refrigerant flow is the same as in the first embodiment. Hereinafter, a method for controlling the first expansion valve 211 will be described. First, Table 1 shows an outline of control in each control mode (full cooling, cooling main, total warming, warm main).
Figure JPOXMLDOC01-appb-T000001
Figure JPOXMLDOC01-appb-T000001
 全ての室内機301~303が冷房運転である全冷房運転(以下、「全冷」と略す)の場合には、第1膨張弁211は、全開状態としておき、室内機301~303の膨張弁321~323のみで負荷に応じた流量制御を行う。 When all the indoor units 301 to 303 are in a cooling only operation (hereinafter, abbreviated as “fully cooled”), the first expansion valve 211 is fully opened, and the expansion valves of the indoor units 301 to 303 Only 321 to 323 performs flow control according to the load.
 図1において、負荷側熱交換器311~313の上側に温度を検知するための温度センサ311a、312a、313aが設けられている。負荷側熱交換器311~313と負荷側の膨張弁321~323の間には、温度を検知するための温度センサ311b、312b、313bと圧力を検知する圧力センサ311c、312c、313cが設けられている。 In FIG. 1, temperature sensors 311a, 312a, and 313a for detecting temperature are provided above the load side heat exchangers 311 to 313. Between the load-side heat exchangers 311 to 313 and the load-side expansion valves 321 to 323, temperature sensors 311b, 312b and 313b for detecting temperature and pressure sensors 311c, 312c and 313c for detecting pressure are provided. ing.
 全冷運転の場合は、温度センサ311a、312a、313aと温度センサ311b、312b、313bとの温度差を演算し、その演算結果を過熱度(スーパーヒート)とし、過熱度が所定の値、例えば2℃程度になるように、負荷側の膨張弁321、322、323の開度を調節する。 In the case of the total cooling operation, the temperature difference between the temperature sensors 311a, 312a, and 313a and the temperature sensors 311b, 312b, and 313b is calculated, and the calculation result is defined as a superheat degree (superheat), and the superheat degree is a predetermined value, for example, The opening degree of the expansion valves 321, 322, and 323 on the load side is adjusted so as to be about 2 ° C.
 冷房サイクルの状態で、冷暖同時運転する「冷主」運転の場合、高圧圧力検知手段281、中間圧圧力検知手段282の差圧を用いた後述する差圧制御を行う。冷主運転においても、負荷側熱交換器が蒸発器として作用しているときは、前述した過熱度を検知しながら負荷側の膨張弁開度を所定の過熱度になるように調整する。 In the case of a “cooling main” operation in which cooling and heating are performed simultaneously in the cooling cycle state, differential pressure control, which will be described later, using the differential pressures of the high pressure detection means 281 and the intermediate pressure detection means 282 is performed. Even in the cold main operation, when the load-side heat exchanger acts as an evaporator, the load-side expansion valve opening is adjusted to a predetermined degree of superheat while detecting the degree of superheat described above.
 また、全ての室内機301~303が暖房運転である全暖房運転(以下、「全暖」と略す)の場合には、第1膨張弁211は、基本的には全閉状態としておき、室内機の膨張弁321~323のみで負荷に応じた流量制御を行い、前記高圧圧力検知手段281、中間圧圧力検知手段282の差圧を用いた後述する差圧制御を行う。 When all the indoor units 301 to 303 are in a heating operation that is a heating operation (hereinafter abbreviated as “fully warm”), the first expansion valve 211 is basically in a fully closed state, The flow rate control according to the load is performed only by the expansion valves 321 to 323 of the machine, and the differential pressure control described later using the differential pressure of the high pressure detecting means 281 and the intermediate pressure detecting means 282 is performed.
 負荷側熱交換器311~313と負荷側の膨張弁321~323の間に、温度を検知するための温度センサ311b、312b、313bと圧力を検知する圧力センサ311c、312c、313cが設けられている。 Between the load-side heat exchangers 311 to 313 and the load-side expansion valves 321 to 323, temperature sensors 311b, 312b and 313b for detecting temperature and pressure sensors 311c, 312c and 313c for detecting pressure are provided. Yes.
 全暖運転の場合は、圧力センサ311c、312c、313cで検知された圧力値から飽和温度Tscを演算する。演算された飽和温度は凝縮温度Tcとする。あらかじめ式(1)で示されるような飽和温度Tscと圧力Pとの関係式を作成しておく必要がある。
   Tsc=f(P)          (1)
In the case of the full warm operation, the saturation temperature T sc is calculated from the pressure values detected by the pressure sensors 311c, 312c, and 313c. The calculated saturation temperature is defined as a condensation temperature Tc . It is necessary to prepare a relational expression between the saturation temperature T sc and the pressure P as shown in the formula (1) in advance.
T sc = f (P) (1)
 なお、後述する図8に示すように、二酸化炭素を冷媒として用いた場合、高圧側は臨界点以上で動作するため、相変化が起こらない。すなわち、飽和温度Tscが存在しない。そこで、冷凍サイクルの実験などを行い、バランスする圧力と吸込み空気温度から仮想の飽和温度を設定する。例えば、圧力100kgf/cm2の時は飽和温度45℃とする。本実施の形態における式(1)は仮想の飽和温度Tscが算出される演算式である。 As shown in FIG. 8 to be described later, when carbon dioxide is used as a refrigerant, the high pressure side operates at a critical point or higher, so that no phase change occurs. That is, there is no saturation temperature T sc . Therefore, an experiment of the refrigeration cycle is performed, and a virtual saturation temperature is set from the pressure to be balanced and the intake air temperature. For example, the saturation temperature is 45 ° C. when the pressure is 100 kgf / cm 2 . Expression (1) in the present embodiment is an arithmetic expression for calculating a virtual saturation temperature T sc .
 圧力センサ311b、312b、313bから得られた圧力値から式(1)を用いて、仮想の凝縮温度Tcを演算する。温度センサ311b、312b、313bから得られた温度TLと仮想凝縮温度Tcとの差(Tc-TL)を求め、この値を過冷却度SC(サブクール)とする。過冷却度SCが所定の値、例えば5℃程度になるように、負荷側の膨張弁321、322、323の開度を調節する。 A virtual condensing temperature T c is calculated from the pressure values obtained from the pressure sensors 311b, 312b, and 313b using the equation (1). A difference (T c −T L ) between the temperature T L obtained from the temperature sensors 311b, 312b, and 313b and the virtual condensing temperature T c is obtained, and this value is set as the degree of supercooling SC (subcool). The opening degree of the expansion valves 321, 322, and 323 on the load side is adjusted so that the degree of supercooling SC becomes a predetermined value, for example, about 5 ° C.
 そして、暖房サイクルの状態で、冷暖同時運転する「暖主」運転の場合も全暖と同様、基本的には全閉状態として、高圧圧力検知手段281、中間圧圧力検知手段282の差圧を用いた後述する差圧制御を行う。 In the case of the “warm main” operation in which the cooling and heating operations are performed simultaneously in the heating cycle state, basically, as in the case of full warming, the fully closed state is basically set, and the differential pressure between the high pressure detection means 281 and the intermediate pressure detection means 282 is changed. The differential pressure control used later is performed.
 次に、制御方法の詳細を、図4及び図5に示す制御フローチャートを用いて説明する。まず、「全冷」運転については、第1膨張弁211は、常に全開状態としておき、室内機の膨張弁321~323のみで負荷に応じた流量制御を行う。 Next, details of the control method will be described with reference to the control flowcharts shown in FIGS. First, in the “fully-cooled” operation, the first expansion valve 211 is always in a fully opened state, and the flow rate control corresponding to the load is performed only by the expansion valves 321 to 323 of the indoor units.
 そして、「冷主」運転については、図4に示すように、圧縮機110の起動などをトリガにして、あらかじめ設定しておいた初期開度L0からスタートし(ステップS41)、スタートから所定時間U経過後(ステップS42)、前記高圧圧力検知手段281と中間圧圧力検知手段282との検出値の差圧ΔPとあらかじめ設定した設定値P1,P2(P1<P2)の比較に応じて、第1膨張弁211の開度を制御する。 For the “cold main” operation, as shown in FIG. 4, starting from the compressor 110 or the like as a trigger, the operation is started from a preset initial opening degree L0 (step S41), and a predetermined time from the start. After a lapse of U (step S42), according to the comparison between the differential pressure ΔP of the detected value of the high pressure detecting means 281 and the intermediate pressure detecting means 282 and the preset set values P1, P2 (P1 <P2) The opening degree of the 1 expansion valve 211 is controlled.
 例えばΔP>P2の場合には、第1膨張弁211をあらかじめ設定した所定開度だけアップし(ステップS43→S44)、P1≦ΔP≦P2の場合には、第1膨張弁211を現状開度に維持し(ステップS43→S45→S46)、P1<ΔPの場合には、第1膨張弁211をあらかじめ設定した所定開度だけダウンさせる(ステップS43→S45→S47→S48)。 For example, when ΔP> P2, the first expansion valve 211 is increased by a predetermined opening degree set in advance (steps S43 → S44), and when P1 ≦ ΔP ≦ P2, the first expansion valve 211 is opened at the current opening degree. (Step S43 → S45 → S46), and when P1 <ΔP, the first expansion valve 211 is lowered by a predetermined opening degree (steps S43 → S45 → S47 → S48).
 以上の制御により、暖房運転室内機側の負荷に応じた冷媒流量を流すだけの必要差圧を確保し、かつ必要以上の差圧をつけてしまうことで、低圧低下を招き、結果的にCOP低下となることを抑制できる。 By the above control, a necessary differential pressure sufficient to flow the refrigerant flow corresponding to the load on the heating operation indoor unit side is secured, and an excessive pressure difference is applied, resulting in a decrease in low pressure, resulting in COP. It can suppress that it falls.
 また、「全暖」運転については、第1膨張弁211は、常に全閉状態としておき、室内機の膨張弁321~323のみで負荷に応じた流量制御を行う。 In the “fully warm” operation, the first expansion valve 211 is always in a fully closed state, and the flow rate is controlled according to the load only with the expansion valves 321 to 323 of the indoor unit.
 そして、「暖主」運転については、図5に示すように、圧縮機110の起動などをトリガにして、第1膨張弁211の全閉状態からスタートし(ステップS51)、スタートから所定時間U経過後(ステップS52)、前記高圧圧力検知手段281と中間圧圧力検知手段282の検出値の差圧ΔPとあらかじめ設定した設定値P1,P2(P1<P2)の比較に応じて、第1膨張弁211の開度を制御する。 Then, as shown in FIG. 5, the “warm main” operation starts from the fully closed state of the first expansion valve 211 with the start of the compressor 110 as a trigger (step S51), and starts a predetermined time U from the start. After the elapse (step S52), the first expansion is performed in accordance with the comparison between the differential pressure ΔP between the detection values of the high pressure detection means 281 and the intermediate pressure detection means 282 and preset values P1, P2 (P1 <P2). The opening degree of the valve 211 is controlled.
 例えば、ΔP>P2の場合には、第1膨張弁211をあらかじめ設定した所定開度だけアップし(ステップS53→S54)、P1≦ΔP≦P2の場合には、第1膨張弁211を現状開度に維持し(ステップS53→S55→S56)、P1<ΔPの場合には、第1膨張弁211をあらかじめ設定した所定開度だけダウンさせる(ステップS53→S55→S57→S58)。 For example, when ΔP> P2, the first expansion valve 211 is increased by a predetermined opening degree set in advance (steps S53 → S54), and when P1 ≦ ΔP ≦ P2, the first expansion valve 211 is currently opened. (Step S53 → S55 → S56), and if P1 <ΔP, the first expansion valve 211 is lowered by a predetermined opening degree (steps S53 → S55 → S57 → S58).
 以上の制御により、暖房運転室内機側の負荷に応じた冷媒流量を流すだけの必要差圧を確保し、かつ必要以上の差圧をつけてしまうことで、冷房運転室内機の入口圧が低圧に近づき、冷房運転室内機で負荷に応じた冷媒流量を流すだけの必要差圧が確保できることになり、結果的にCOP低下となることを抑制できる。また、圧力の変化を抑制し、室内機へ安定して冷媒を送り込むことができ、省エネ運転と快適性を実現することができる。 The above control ensures the necessary differential pressure to flow the refrigerant flow according to the load on the heating operation indoor unit side, and applies the differential pressure more than necessary, thereby reducing the inlet pressure of the cooling operation indoor unit. As a result, the necessary differential pressure sufficient to allow the refrigerant flow rate corresponding to the load to flow in the cooling operation indoor unit can be secured, and as a result, it is possible to suppress a COP decrease. Moreover, a change in pressure can be suppressed, and the refrigerant can be stably sent to the indoor unit, so that energy saving operation and comfort can be realized.
 実施の形態3.
 この実施の形態3では、前述した実施の形態2と同様に、図1及び図3に示す実施の形態1の構成と同様な構成を備える。さらに、図1及び図3において、分流コントローラ200内に、実施の形態1では不要であった、高圧圧力検知手段281、中間圧圧力検知手段282、第1温度検知手段291、第2温度検知手段292が備えられる。
Embodiment 3 FIG.
In the third embodiment, the same configuration as the configuration of the first embodiment shown in FIG. 1 and FIG. Further, in FIG. 1 and FIG. 3, the high pressure detection means 281, the intermediate pressure detection means 282, the first temperature detection means 291, and the second temperature detection means that are not required in the first embodiment are included in the shunt controller 200. 292 is provided.
 この実施の形態3において、冷媒の流れは、実施の形態1と同じである。以下、第2膨張弁212の制御方法について説明する。まず、各制御モード(全冷、冷主、全暖、暖主)における制御概要は表1の通りである。 In the third embodiment, the refrigerant flow is the same as in the first embodiment. Hereinafter, a method for controlling the second expansion valve 212 will be described. First, Table 1 shows an outline of control in each control mode (full cooling, cool main, full warm, warm main).
 全ての室内機301~303が冷房運転である「全冷」運転の場合には、第2膨張弁212は、第1温度検知手段291、第2温度検知手段292を用いた後述する差温(スーパーヒート)制御を行い、室内機301~303側では膨張弁321~323にて負荷に応じた流量制御を行う。冷房サイクルの状態で、冷暖同時運転する「冷主」運転の場合も「全冷」と同様である。 When all the indoor units 301 to 303 are in the “cooling” operation, which is a cooling operation, the second expansion valve 212 uses the first temperature detection means 291 and the second temperature detection means 292, which will be described later. Superheat) control is performed, and on the indoor units 301 to 303 side, the expansion valves 321 to 323 perform flow rate control according to the load. The “cooling main” operation in which the cooling and heating operations are performed simultaneously in the cooling cycle state is the same as “full cooling”.
 また、全ての室内機301~303が暖房運転である「全暖」運転の場合には、第2膨張弁212は、全開状態としておき、室内機の膨張弁321~323で負荷に応じた流量制御を行い、負荷側と熱交換した冷媒はこの第2膨張弁212を介して、室外機低圧ラインへ流入する。 When all the indoor units 301 to 303 are in the “fully warm” operation, which is a heating operation, the second expansion valve 212 is kept fully open, and the flow rate corresponding to the load is set by the expansion valves 321 to 323 of the indoor units. The refrigerant that has been controlled and exchanged heat with the load side flows into the outdoor unit low-pressure line via the second expansion valve 212.
 そして、暖房サイクルの状態で、冷暖同時運転する「暖主」運転の場合は、基本的には全閉状態として、高圧圧力検知手段281、中間圧圧力検知手段282の差圧を用いた後述する差圧制御を行う。 In the case of the “warm main” operation in which the cooling and heating operation is performed simultaneously in the heating cycle state, basically, as will be described later, the differential pressure between the high pressure detection means 281 and the intermediate pressure detection means 282 is used as the fully closed state. Perform differential pressure control.
 次に、制御方法の詳細を図6及び図7に示す制御フローチャートを用いて説明する。まず、「全冷」運転については、図6に示すように、圧縮機110の起動などをトリガにして、あらかじめ設定しておいた初期開度L0からスタートし(ステップS61)、スタートから所定時間U経過後(ステップS62)、前記第1温度検知手段291、第2温度検知手段292を用いてその検出値の差温ΔT(スーパーヒート)を演算し、その差温ΔTとあらかじめ設定したT1,T2(T1<T2)との比較に応じて、第2膨張弁212の開度を制御する。 Next, details of the control method will be described with reference to control flowcharts shown in FIGS. First, as shown in FIG. 6, the “all-cooling” operation starts from the initial opening L0 set in advance by using the start of the compressor 110 as a trigger (step S61), and starts for a predetermined time from the start. After the elapse of U (step S62), the temperature difference ΔT (superheat) of the detected values is calculated using the first temperature detecting means 291 and the second temperature detecting means 292, and the temperature difference ΔT is set to T1, which is set in advance. The opening degree of the second expansion valve 212 is controlled according to the comparison with T2 (T1 <T2).
 例えば、ΔT>T2の場合には、第2膨張弁212をあらかじめ設定した所定開度だけアップし(ステップS63→S64)、T1≦ΔT≦T2の場合には、第2膨張弁212を現状開度に維持し(ステップS63→S65→S66)、T1<ΔTの場合には、第2膨張弁212をあらかじめ設定した所定開度だけダウンさせる(ステップS63→S65→S67→S68)。 For example, when ΔT> T2, the second expansion valve 212 is increased by a predetermined opening degree set in advance (steps S63 → S64), and when T1 ≦ ΔT ≦ T2, the second expansion valve 212 is currently opened. (Step S63 → S65 → S66), and when T1 <ΔT, the second expansion valve 212 is lowered by a predetermined opening degree (steps S63 → S65 → S67 → S68).
 以上の制御により、冷房運転室内機側入口の冷媒温度を冷却し、性能を満足するだけの必要エンタルピ差を確保し、結果的にCOP低下となることを抑制できる。また、冷房主体で一部暖房運転となる冷房主体運転においても、冷房運転室内機へより低温の冷媒を送り込むことができ、省エネ運転を実現することができる。 With the above control, the refrigerant temperature at the cooling operation indoor unit side inlet is cooled, and a necessary enthalpy difference sufficient to satisfy the performance can be secured, so that it is possible to suppress the decrease in COP. Further, even in the cooling main operation in which the cooling main operation is a partial heating operation, a lower temperature refrigerant can be sent to the cooling operation indoor unit, and an energy saving operation can be realized.
 また、「全暖」運転については、第2膨張弁212は、常に全開状態としておき、室内機の膨張弁321~323で負荷に応じた流量制御を行い、負荷側と熱交換した冷媒はこの第2膨張弁212を介して室外機低圧ラインへ流入する。 In the “fully warm” operation, the second expansion valve 212 is always fully opened, the flow rate is controlled according to the load by the expansion valves 321 to 323 of the indoor units, and the refrigerant exchanging heat with the load side is It flows into the outdoor unit low pressure line through the second expansion valve 212.
 そして、「暖主」運転については、図7に示すように、圧縮機110の起動などをトリガにして、全閉状態からスタートし(ステップS71)、スタートから所定時間U経過後(ステップS72)、前記高圧圧力検知手段281、中間圧圧力検知手段282の検出値の差圧ΔPとあらかじめ設定した設定値P1,P2(P1<P2)との比較に応じて、第2膨張弁212の開度を制御する。 As shown in FIG. 7, the “warm main” operation starts from the fully closed state triggered by the start of the compressor 110 (step S71), and after a predetermined time U has elapsed from the start (step S72). The opening degree of the second expansion valve 212 is determined in accordance with a comparison between the pressure difference ΔP detected by the high pressure detection means 281 and the intermediate pressure detection means 282 and preset values P1, P2 (P1 <P2). To control.
 例えば、ΔP>P2の場合には、第2膨張弁212をあらかじめ設定した所定開度だけダウンし(ステップS73→S74)、P1≦ΔP≦P2の場合には、第2膨張弁212を現状開度に維持し(ステップS73→S75→S76)、P1<ΔPの場合には、第2膨張弁212をあらかじめ設定した所定開度だけアップさせる(ステップS73→S75→S77→S78)。 For example, if ΔP> P2, the second expansion valve 212 is lowered by a predetermined opening degree set in advance (steps S73 → S74), and if P1 ≦ ΔP ≦ P2, the second expansion valve 212 is currently opened. (Step S73 → S75 → S76), and when P1 <ΔP, the second expansion valve 212 is increased by a predetermined opening degree (steps S73 → S75 → S77 → S78).
 以上の制御により、暖房運転室内機側の負荷に応じた冷媒流量を流すだけの必要差圧を確保し、かつ必要以上の差圧をつけてしまう(中間圧が低圧に近づく)ことで、冷房運転室内機の入口圧が低圧に近づき、冷房運転室内機で負荷に応じた冷媒流量を流すだけの必要差圧が確保できなくなり、結果的にCOP低下となることを抑制できる。 By the above control, the necessary differential pressure that allows the refrigerant flow rate to flow according to the load on the heating operation indoor unit side is secured, and the differential pressure more than necessary is applied (the intermediate pressure approaches low pressure), thereby cooling the system. The inlet pressure of the operating indoor unit approaches a low pressure, and the required differential pressure sufficient to flow the refrigerant flow rate corresponding to the load cannot be secured in the cooling operation indoor unit, and as a result, it is possible to suppress a COP decrease.
 冷主運転時においては、熱源機側熱交換器130は凝縮器(放熱器)と作用する。冷主運転は冷房負荷が暖房負荷を上回っている状態なので、放熱能力の一部を熱源機側熱交換器130で放熱する必要がある。そのためには、ファン速度や熱源機側熱交換器130を分割し、熱交換器容量を増減させる必要がある。 During the cold main operation, the heat source machine side heat exchanger 130 acts with a condenser (heat radiator). In the cooling main operation, since the cooling load is higher than the heating load, it is necessary to radiate a part of the heat radiation capability with the heat source unit side heat exchanger 130. For that purpose, it is necessary to divide the fan speed and the heat source device side heat exchanger 130 to increase or decrease the heat exchanger capacity.
 この発明では、熱源機側熱交換器130を分割することなく、放熱能力を調整する方法を説明する。本実施の形態では、冷媒として二酸化炭素を用いている。この冷媒は図2や図8で示すように、臨界点以上で高圧側が動作する。この特性を利用すれば容易に放熱容量を調整することが可能となる。 In the present invention, a method for adjusting the heat radiation capacity without dividing the heat source apparatus side heat exchanger 130 will be described. In the present embodiment, carbon dioxide is used as the refrigerant. As shown in FIGS. 2 and 8, this refrigerant operates on the high pressure side above the critical point. If this characteristic is used, the heat radiation capacity can be easily adjusted.
 図3に示すように、熱源機側熱交換器130と逆止弁141との間に、圧力センサ900、温度センサ901を設けると共に、熱源機側熱交換器130の入口に温度センサ902を設ける。超臨界では温度と圧力が決まれば、一義的にエンタルピが決定される。 As shown in FIG. 3, a pressure sensor 900 and a temperature sensor 901 are provided between the heat source device side heat exchanger 130 and the check valve 141, and a temperature sensor 902 is provided at the inlet of the heat source device side heat exchanger 130. . In supercritical, if temperature and pressure are determined, enthalpy is uniquely determined.
 図8において、aは熱源機側熱交換器130入口のエンタルピH1、bは熱源機側熱交換器130出口のエンタルピH2(暖房運転になっている負荷側熱交換器の入口)、cは熱源機側熱交換器130入口のエンタルピH3である。 In FIG. 8, a is the enthalpy H 1 at the inlet of the heat source machine side heat exchanger 130, b is the enthalpy H 2 at the outlet of the heat source machine side heat exchanger 130 (inlet of the load side heat exchanger in the heating operation), c Is the enthalpy H 3 at the inlet of the heat source machine side heat exchanger 130.
 暖房側の負荷は暖房運転を行っている台数と接続されている室内機の容量から知ることができる。この暖房負荷をQcとする。また、圧縮機の吐出圧力、吸入圧力から冷媒流量Grを求めることができる。式(2)から暖房負荷を賄うのに必要なエンタルピ差ΔHを求めることができる。
   Qc/Gr=ΔH=H2-H3     (2)
The load on the heating side can be determined from the number of heating operations and the capacity of the connected indoor units. Let this heating load be Q c . Further, the refrigerant flow rate G r can be obtained from the discharge pressure and suction pressure of the compressor. The enthalpy difference ΔH necessary to cover the heating load can be obtained from the equation (2).
Q c / G r = ΔH = H 2 −H 3 (2)
 また、負荷側熱交換器出口の圧力センサ311cと温度センサ311bから負荷側熱交換器(暖房)出口のエンタルピH3を求めることができる。式(2)は式(3)のように書き換えられる。
   H2=Qc/Gr+H3        (3)
Further, the enthalpy H 3 at the load side heat exchanger (heating) outlet can be obtained from the pressure sensor 311c and the temperature sensor 311b at the load side heat exchanger outlet. Equation (2) can be rewritten as Equation (3).
H 2 = Q c / G r + H 3 (3)
 すなわち、熱源機側熱交換器出口のエンタルピH2が決定される。熱源機側熱交換器出口のエンタルピは圧力センサ900、温度センサ901から求めることができる。式(3)から求められたエンタルピを目標のエンタルピH2mとする。 That is, the enthalpy H 2 at the outlet of the heat source device side heat exchanger is determined. The enthalpy at the heat source side heat exchanger outlet can be obtained from the pressure sensor 900 and the temperature sensor 901. The enthalpy obtained from the equation (3) is set as the target enthalpy H 2m .
 また、圧力センサ900、温度センサ901で測定されたエンタルピをH2sとする。目標エンタルピH2mと測定されたエンタルピH2sの差(H2m-H2s)を演算し、H2m-H2s=ΔHsとする。 Further, the enthalpy measured by the pressure sensor 900 and the temperature sensor 901 is assumed to be H 2s . The difference (H 2m -H 2s ) between the target enthalpy H 2m and the measured enthalpy H 2s is calculated to be H 2m -H 2s = ΔH s .
 そして、-epsH2<ΔHs<epsH2(なお、-epsH2とepsH2は誤差範囲の下限値と上限値を示す)が所定の値になるように制御手段により熱源機側ファン(送風機)の回転を増減させる。-epsH2>ΔHsの場合は、ファン回転数をアップさせ、ΔHs>epsH2の場合はファン回転数を小さくする制御を行う。 And, -epsH 2 <ΔH s <epsH 2 (where -epsH 2 and epsH 2 indicate the lower limit value and the upper limit value of the error range) to be a predetermined value by the control means by the heat source side fan (blower) Increase or decrease the rotation of. When -epsH 2 > ΔH s , the fan speed is increased, and when ΔH s > epsH 2 , the fan speed is decreased.
 なお、エンタルピHを求めるためには、あらかじめ式(4)の物性式を作成しておく必要ある。
   H=f(P,T)          (4)
   P:圧力、T:温度
In addition, in order to obtain enthalpy H, it is necessary to create a physical property formula of Formula (4) in advance.
H = f (P, T) (4)
P: Pressure, T: Temperature

Claims (6)

  1.  室外機と複数の室内機とを分流コントローラにより接続し、超臨界流体を用いて1つの冷凍サイクルを構成した空気調和システムにおいて、
     前記室外機と前記分流コントローラとの間を、高圧配管及び低圧配管の2管で接続し、
     前記分流コントローラと前記複数の室内機との間を、高圧配管及び低圧配管の2管で接続し、
     前記分流コントローラは、
     前記室外機から前記室内機への冷媒を分岐して第1膨張弁により減圧した冷媒と前記室内機からの冷媒とが合流して流入される比較的高温の中間圧二相冷媒と、前記室内機へ流出される冷媒を分岐して第2膨張弁により減圧して前記室外機へ流出される比較的低温の低圧二相冷媒とを熱交換させる二重管熱交換器を有する
     ことを特徴とする空気調和装置。
    In an air conditioning system in which an outdoor unit and a plurality of indoor units are connected by a shunt controller and one refrigeration cycle is configured using a supercritical fluid,
    Between the outdoor unit and the diversion controller are connected by two pipes, a high pressure pipe and a low pressure pipe,
    Between the shunt controller and the plurality of indoor units is connected by two pipes, a high pressure pipe and a low pressure pipe,
    The shunt controller is
    A relatively high-temperature intermediate-pressure two-phase refrigerant into which the refrigerant branched from the outdoor unit to the indoor unit and depressurized by the first expansion valve and the refrigerant from the indoor unit are combined and flowed; A double pipe heat exchanger for branching the refrigerant flowing out to the machine, reducing the pressure by the second expansion valve, and exchanging heat with the relatively low-temperature low-pressure two-phase refrigerant flowing out to the outdoor unit, Air conditioner to do.
  2.  請求項1に記載の空気調和装置において、
     前記二重管熱交換器は、プレート熱交換器またはマイクロチャネル熱交換器でなる
     ことを特徴とする空気調和装置。
    In the air conditioning apparatus according to claim 1,
    The air conditioner characterized in that the double pipe heat exchanger is a plate heat exchanger or a microchannel heat exchanger.
  3.  請求項1に記載の空気調和装置において、
     前記第1膨張弁の出入口に、それぞれ圧力検知手段を設け、
     前記第1膨張弁の開度は、前記圧力検知手段の2つの圧力検知値の差圧が一定となるように制御される
     ことを特徴とする空気調和装置。
    In the air conditioning apparatus according to claim 1,
    Pressure detecting means are provided at the inlet and outlet of the first expansion valve,
    The opening degree of the first expansion valve is controlled such that the differential pressure between the two pressure detection values of the pressure detection means is constant.
  4.  請求項1に記載の空気調和装置において、
     前記第2膨張弁の出口と、前記二重管熱交換器の低圧側出口とに、それぞれ温度検出手段を設け、
     前記第2膨張弁の開度は、前記温度検出手段の2つの温度検知値の差温が一定となるように制御される
     ことを特徴とする空気調和装置。
    In the air conditioning apparatus according to claim 1,
    Temperature detection means are provided at the outlet of the second expansion valve and the low pressure side outlet of the double pipe heat exchanger,
    The opening degree of the second expansion valve is controlled so that the difference between the two temperature detection values of the temperature detection means is constant.
  5.  請求項1に記載の空気調和装置において、
     前記第1膨張弁の出入口に、それぞれ圧力検知手段を設け、
     前記第2膨張弁の開度は、前記圧力検知手段の2つの圧力検知値の差圧が一定となるように制御される
     ことを特徴とする空気調和装置。
    In the air conditioning apparatus according to claim 1,
    Pressure detecting means are provided at the inlet and outlet of the first expansion valve,
    The opening degree of the second expansion valve is controlled such that the differential pressure between the two pressure detection values of the pressure detection means is constant.
  6.  請求項1に記載の空気調和装置において、
     熱源機側熱交換器の出口に、圧力検知手段及び温度検知手段と、前記圧力検知手段の値と前記温度検知手段の値とからエンタルピを演算する制御手段とを備え、目標のエンタルピになるように熱源機側の送風機の回転速度を増減させる
     ことを特徴とする空気調和装置。
    In the air conditioning apparatus according to claim 1,
    At the outlet of the heat source machine side heat exchanger, a pressure detection means and a temperature detection means, and a control means for calculating an enthalpy from the value of the pressure detection means and the value of the temperature detection means are provided so as to achieve a target enthalpy. The air conditioner characterized by increasing or decreasing the rotational speed of the fan on the heat source side.
PCT/JP2009/056655 2008-03-31 2009-03-31 Air conditioner WO2009123190A1 (en)

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