WO2006013970A1 - Freezing cycle apparatus - Google Patents

Freezing cycle apparatus Download PDF

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Publication number
WO2006013970A1
WO2006013970A1 PCT/JP2005/014416 JP2005014416W WO2006013970A1 WO 2006013970 A1 WO2006013970 A1 WO 2006013970A1 JP 2005014416 W JP2005014416 W JP 2005014416W WO 2006013970 A1 WO2006013970 A1 WO 2006013970A1
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WO
WIPO (PCT)
Prior art keywords
pressure
low
stage compressor
refrigerant
expander
Prior art date
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PCT/JP2005/014416
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French (fr)
Japanese (ja)
Inventor
Takahiro Yamaguchi
Shuuji Fujimoto
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Daikin Industries, Ltd.
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Application filed by Daikin Industries, Ltd. filed Critical Daikin Industries, Ltd.
Publication of WO2006013970A1 publication Critical patent/WO2006013970A1/en

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • F25B1/10—Compression machines, plants or systems with non-reversible cycle with multi-stage compression
    • 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
    • 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

Definitions

  • the present invention relates to a refrigeration cycle apparatus including an expander for energy recovery.
  • Japanese Patent Laid-Open No. 2001-116371 discloses the following energy recovery type refrigeration cycle apparatus in the prior art.
  • a positive displacement expander that performs substantially isentropic expansion provided in a general refrigerant circuit is arranged coaxially with the compressor, and the energy recovered by the expander is used as power for the compressor.
  • the mass circulation amount of the refrigerant flowing through the compressor and the mass circulation amount of the refrigerant circulating through the expander must be equal, while the volume circulation amount passing through both is respectively Since it is determined by “cylinder volume X rotation speed”, the restriction of the constant density ratio shown in the following equation is imposed.
  • VC Volume circulation volume of refrigerant sucked into the compressor
  • VC and VE are values specific to the refrigeration cycle system, so VCZVE is It is a constant value.
  • Japanese Patent Laid-Open No. 2001-116371 solves this problem in the prior art.
  • a bypass circuit having a flow control valve interposed is provided in parallel with the expander. Then, by adjusting the flow rate of refrigerant flowing through the bypass circuit in response to changes in operating conditions, the flow rate of refrigerant passing through the expander is adjusted, and the constant density ratio and the restrictions on the refrigeration cycle are eliminated. To improve efficiency.
  • the refrigeration cycle apparatus has been made to solve the above-mentioned problems, and the first invention is a two-stage compression apparatus comprising a low-stage compressor and a high-stage compressor.
  • a decompressor including a radiator that cools the high-pressure gas refrigerant, an expander that expands the high-pressure refrigerant after being cooled by the radiator, an evaporator that evaporates the low-pressure refrigerant decompressed by the decompressor, and a decompression
  • a gas injection circuit that introduces the intermediate-pressure gas refrigerant decompressed by the apparatus to the suction side of the high-stage compression, and the expander and the low-stage compressor are connected coaxially.
  • the gas injection circuit for introducing the intermediate-pressure gas refrigerant to the suction side of the high-stage compressor is provided, the energy recovered in the expander that depressurizes the high-pressure gas refrigerant as compared with the prior arts 1 and 2. Although it decreases, the compression work in the two-stage compressor decreases. Further, the energy recovery efficiency in the prior art is not so great. Due to such factors, the present invention can improve the energy efficiency by reducing the energy consumption in the compressor.
  • FIG. 1 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 1.
  • FIG. 2 is a Mollier diagram of the refrigeration cycle apparatus.
  • FIG. 3 is a refrigeration (or cooling) COP ratio diagram in the refrigeration cycle apparatus according to Embodiment 1.
  • FIG. 4 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 2.
  • FIG. 5 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 3.
  • FIG. 6 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 4.
  • FIG. 7 is a Mollier diagram of the refrigeration cycle apparatus.
  • FIG. 8 is a hot water supply (or heating) COP ratio diagram in the refrigeration cycle apparatus according to Embodiment 5.
  • FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 6.
  • the refrigeration cycle apparatus includes a two-stage compressor 3 composed of a low-stage compressor 1 and a high-stage compressor 2.
  • the high-pressure gas refrigerant released from the compression device 3 is cooled by the radiator 4 and then separated by the gas-liquid separator 6 as an intermediate-pressure refrigerant decompressed by the expander 5.
  • the liquid separated from the gas by the gas-liquid separator 6 is depressurized to a low pressure by the expansion device 7 and then evaporated and evaporated by the evaporator 8 and sent to the suction port of the low-stage compressor 1 via the accumulator 9. .
  • the gas separated from the refrigerant liquid in the gas-liquid separator 6 is sent to the suction port of the high-stage compressor 2 through the gas injection circuit 10.
  • the low-stage and high-stage compressors 1 and 2 are constant capacity compressors such as a rotary compressor.
  • the low stage compressor 1 and the expander 5 are connected coaxially.
  • the flow rate of the circuit 10 may be appropriately set by adjusting the opening degree of the on-off valve 12 of the gas instruction circuit 10.
  • the opening degree of the flow control valve 13 of the bypass circuit 11 is controlled by the suction pressure of the low-stage compressor 1.
  • the flow control valve 13 is controlled so that the opening degree of the refrigerant is increased when the pressure of the refrigerant sucked into the low-stage compressor 1 increases.
  • the expander 5 and the expansion device 7 constitute a decompression device in the present invention.
  • the refrigeration cycle apparatus configured in this manner is filled with a refrigerant that forms a normal refrigeration cycle such as HFC32 or HFC32 (a refrigerant that does not form a supercritical refrigeration cycle) as a refrigerant.
  • the refrigerant (a5) decompressed to the intermediate pressure is gas-liquid separated by the gas-liquid separator 6.
  • the separated gas refrigerant (a8) is introduced to the suction side of the high-stage compressor 2 as described above and mixed with the discharge gas (al) of the low-stage compressor 1 (a2).
  • the liquid refrigerant (a7) separated by the gas-liquid separator 6 is decompressed by the expansion device 7 to become a low-pressure refrigerant (a8).
  • This low-pressure refrigerant is evaporated and evaporated by the evaporator 8 (a9), and is sucked into the low-stage compressor 1 through the accumulator 9.
  • the accumulator 9 separates and stores the liquid refrigerant that has not evaporated in the suction side of the low-stage compressor 1 to store the liquid refrigerant in the low-stage compressor 1. This is to prevent this.
  • the opening degree of the flow control valve 13 of the bypass circuit 11 is minimized (that is, (Closed state) to satisfy the constraint of a constant density ratio.
  • the flow control valve 13 corresponds to the increase. Control is performed to increase the amount of refrigerant circulation in the bypass circuit 11 by increasing the opening.
  • surplus refrigerant that does not pass through the expander 5 among the high-pressure refrigerant cooled by the radiator 4 is caused to flow to the gas-liquid separator 6 through the bypass circuit 11.
  • the high-pressure refrigerant is reduced to an intermediate pressure by the flow control valve 13 (a6).
  • it is mixed with the intermediate-pressure refrigerant that has flowed to the gas-liquid separator 6 via the expander 5 described above, and the gas-liquid separator 6 performs gas-liquid separation.
  • the density is controlled by providing the no-pass circuit 11 including the flow control valve 13 in parallel with the expander 5 and controlling the amount of bypass to increase in response to the increase in the evaporation temperature. Since the constant constraint force is also released, it is possible to prevent the Mollier diagram from becoming an inefficient vertical diagram, and to improve efficiency.
  • FIG. 3 shows that when the refrigeration cycle apparatus of Embodiment 1 is a refrigeration apparatus (or a cooling apparatus), the refrigeration (or cooling) COP ratio corresponding to the change in the evaporation temperature can be compared with the prior art. It shows.
  • the refrigeration (or cooling) COP ratio in this case refers to the refrigeration (or cooling) COP at each evaporating temperature of an old general refrigeration apparatus (or cooling apparatus) that does not perform energy recovery by the expander 5. This is expressed as the ratio of COP in the first and second embodiments.
  • the low-stage compressor 1 of the first embodiment is of a variable capacity type, and a bypass circuit 11 having a flow control valve 13 and an on-off valve 12 in the gas injection circuit 10 are provided. It is omitted. Further, the low-stage compressor 1 and the expander 5 are accommodated in the same casing 21. That is, the casing 21 is divided into two chambers by the partition wall 22, the low-stage compressor 1 is stored in one chamber 23, and the expander 5 is stored in the other chamber 24. The outlet port of the accumulator 9 is connected to the compressor chamber 23, and the low-stage compressor 1 sucks the refrigerant in the compressor chamber 23.
  • the Mollier diagram in this refrigeration cycle apparatus is partially different from FIG.
  • the Mollier diagram is also a diagram in which the broken line portion showing the substantially isenthalpy change (a4 ⁇ a6) by the flow control valve 13 is omitted in FIG. It becomes.
  • a variable capacity swash plate compressor is used as the variable capacity low stage compressor 1 described above.
  • variable-capacity low-stage compressor 1 is controlled as follows. That is, when the evaporation temperature is the lowest temperature in the allowable operating range (for example, ⁇ 35 ° C.), the rotation speed of the low-stage compressor 1 is set to be the maximum. When the evaporation load of the evaporator 8 increases, the suction pressure of the low-stage compressor 1 increases. The capacity of the low-stage compressor 1 is controlled to decrease in response to the increase in the suction pressure of the high-stage compressor 2 accompanying this.
  • the low-stage compressor 1 is used as a variable displacement compressor, and the capacity of the low-stage compressor 1 is controlled in response to changes in operating conditions.
  • the Mollier diagram can be prevented from becoming an inefficient vertical diagram.
  • the capacity of the low-stage compressor 1 as compared to the case where the expansion unit is bypassed and the constant density ratio is released as in the case of the first embodiment 1, The efficiency can be further improved.
  • the low-stage compressor 1 is a variable capacity swash plate compressor, the density ratio of the refrigerant flowing into the expander 5 and the density ratio of the refrigerant flowing into the high-stage compressor 2 And adjust steplessly Energy efficiency can be further improved.
  • the low-stage compressor 1, the gas-liquid separator 6, and the accumulator 9 of the second embodiment are integrated. That is, in the third embodiment, the lower part of the compressor chamber 23 is an accumulator.
  • the outlet port of the evaporator 8 is directly connected to the compressor chamber 23, and the low-stage compressor 1 directly sucks the refrigerant in the compressor chamber 23. Therefore, the liquid refrigerant returning from the evaporator 8 is stored in the lower part of the compressor chamber 23, so that the compressor chamber 23 also functions as an accumulator.
  • the expander chamber 24 is also used as a gas-liquid separator.
  • the outlet port of the expander 5 is opened in the expander chamber 24, and a part of the expander chamber 24 is connected to the evaporator 8 via the expansion device 7.
  • the intermediate pressure refrigerant that has passed through the expander 5 is discharged into the compressor chamber 23.
  • the intermediate-pressure refrigerant in the gas-liquid mixture is separated into gas and liquid by the difference in gravity in the space of the compressor chamber 23. Further, the separated liquid refrigerant flows out to the evaporator 8 through the expansion device 7.
  • the Mollier diagram of this embodiment is the same as that of Embodiment 2, and the energy efficiency with respect to the change in evaporation temperature is shown in FIG.
  • the stage side compressor 1 and the expander 5 are accommodated in the same casing 21 partitioned by the partition wall 22, and further, the low stage side compressor 1 is accommodated.
  • the compressor chamber 23 is also used as an accumulator. Therefore, in the refrigeration cycle apparatus according to the third embodiment, the apparatus is more compact and the accumulator can be installed in other units as compared to the case where the low-stage compressor 1, the expander 5 and the accumulator are individually manufactured and installed. Piping connected to equipment is simplified.
  • the expander chamber 24 containing the expander 5 is used as a gas-liquid separator, a low-stage compressor, an expander, a gas-liquid separator, and an accumulator are individually manufactured and attached. Compared to the refrigeration cycle device, the equipment is compact and the piping connecting these equipment is simplified.
  • the refrigeration cycle equipment consists of a two-stage compression device 3 consisting of a low-stage compressor 1 and a high-stage compressor 2, a radiator 4, a throttling device 31, a throttling device 31, an intermediate-pressure refrigerant and the remaining high-pressure refrigerant.
  • a heat exchanger 32 for exchanging heat with each other, an expander 33 for reducing the intermediate pressure refrigerant cooled by the heat exchanger 32 to a low pressure, an evaporator 8 and an accumulator 9 are configured.
  • the heat exchange 32 is connected to the high-pressure side path 32a through which the high-pressure refrigerant flows and the intermediate pressure-side path 32b through which the intermediate pressure refrigerant flows. And are arranged close to each other.
  • the inlet of the high-pressure side path 32 a of the heat exchanger 32 is connected to the outlet of the radiator 4, and the outlet of the high-pressure side path 32 a is connected to the expander 33.
  • the inlet of the intermediate pressure side passage 32b is connected to the expansion device 31, and the outlet is connected to the suction port of the high stage compressor 2 by the gas injection circuit 34.
  • the low-stage compressor 1 is a variable capacity swash plate compressor
  • the high-stage compressor 2 is a constant capacity compressor such as a single compressor.
  • the low-stage compressor 1 and the expander 33 are connected to the same shaft. It shall be filled with a refrigerant that forms a normal refrigeration cycle such as HFC32 or HFC32 (a refrigerant that does not form a supercritical refrigeration cycle).
  • radiator 4 (b4) A part of the high-pressure refrigerant (b4) cooled by the radiator 4 is decompressed to an intermediate pressure by the throttling device 31 under a pressure-reducing action of a substantially isenthalpy change (b
  • the decompressed refrigerant (b8) is vaporized by exchanging heat with the remaining high-pressure refrigerant by heat exchange (b
  • the high-pressure refrigerant (b5) cooled by the heat exchanger 32 is decompressed to a low pressure by undergoing a substantially isentropic change by the expander 33 (b6).
  • the energy recovered by the expander 33 is used as energy for rotating the low-stage compressor 1.
  • the low-pressure refrigerant (b6) evaporates and evaporates in the evaporator 8 (b7) and is sucked into the low-stage compressor 1.
  • the variable capacity low stage compressor 1 is controlled in the same manner as in the second embodiment.
  • the evaporation temperature is set to the lowest temperature within the allowable operating range (for example, -35 ° C)
  • the constant density ratio is restricted.
  • the evaporation load of the evaporator 8 increases and the suction pressure of the high-stage compressor 2 increases, the capacity of the low-stage compressor 1 is reduced corresponding to the increase.
  • the low-stage compressor 1 is a variable displacement swash plate compressor, and therefore, the density ratio on the suction side of the compressor is adjusted in response to changes in operating conditions. This prevents the Mollier diagram from becoming an inefficient vertical diagram.
  • the expander 33 expands the remaining high-pressure refrigerant after being cooled by the radiator 4 and the heat exchanger 32 to a low pressure between the high-pressure side path of the heat exchanger 32 and the evaporator 8. Is provided. Accordingly, energy corresponding to the decompression of the refrigerant is recovered by the expander 33. In this case, the energy recovered by the expander 33 is circulated to the high-stage compressor 2 through the gas injection circuit 34 as an intermediate pressure gas refrigerant. Accordingly, the amount of compression work in the low-stage compressor 1 is reduced by the amount of refrigerant circulating introduced into the high-stage compressor 2 via the heat exchanger 32, so that energy efficiency is improved.
  • This embodiment is a refrigeration cycle apparatus using carbon dioxide as a refrigerant in the refrigerant circuit of the first embodiment.
  • water is heated in the radiator 4 and the hot water obtained by this is used for hot water supply or heating.
  • the Mollier diagram in this case is a force supercritical cycle in the cycle as shown in Fig. 2, and the high-pressure refrigerant discharged from the high-stage compressor 2 outlet and the high-pressure refrigerant cooled by the radiator 4 are These are high-pressure gas refrigerants having a pressure higher than the critical point. Therefore, there is no room for liquid refrigerant in the high-pressure circuit, and the amount of refrigerant is adjusted by the accumulator 9.
  • the size of the accumulator 9 is set to a size that allows excess refrigerant to be stored.
  • FIG. 8 shows the effect of the hot water supply (or heating) COP ratio on the change in hot water temperature at the outlet of the radiator 4 (indirectly, change in high pressure) in the refrigeration cycle apparatus of the fifth embodiment.
  • the hot water (or heating) COP ratio in this case is defined as the hot water temperature at the condenser outlet of a hot water supply device (or heating device) that applies an old general refrigeration cycle device that does not recover energy by the expander 5.
  • FIG. 7 shows the ratio of the COP of Embodiment 5 to the hot water supply (or heating) COP when changed.
  • the hot water temperature at the radiator 4 outlet is low (that is, when the high pressure is low)
  • the suction pressure of the high-stage compressor 2 is also low, and the amount of refrigerant bypassed from the bypass circuit 11 is reduced. Therefore, the amount of energy recovered by the expander 5 increases, and the hot water supply (or heating) COP ratio increases.
  • the amount of reduction in compression energy in the two-stage compressor 3 is greater than the reduction in recovered energy by the expander 5.
  • the hot water supply (or heating) COP ratio increases.
  • a refrigeration cycle apparatus in the same refrigerant circuit as in the first embodiment, a non-displacement turbine type expander is used as the expander 5, and a non-displacement turbine type compressor is used as the low-stage compressor 1. is doing.
  • carbon dioxide can be used as a refrigerant as in the fifth embodiment.
  • the refrigeration cycle is a supercritical cycle.
  • the power using a variable capacity swash plate type compressor as the variable capacity low stage side compressor 1 is not limited to this, but is an inverter driven rotary compressor, etc. It can also be used as a compressor of other types using other capacity variable methods.
  • the force with which the low-stage compressor 1, the accumulator 9, the expander 5 and the gas-liquid separator 6 are combined is required to be integrated with each other. Absent. Any one of these can be integrated to simplify the equipment configuration and equipment connection piping.
  • the low stage compressor 1 is not a variable capacity compressor, but a fixed capacity compressor, and in parallel with the expander 33 as in the first embodiment, the flow rate A bypass circuit having a control valve may be provided to control the opening degree of the flow control valve to increase as the suction pressure of the high-stage compressor increases.
  • expansion devices 7 and 31 various devices such as an electric expansion valve, a capillary tube, and a temperature-sensitive expansion valve can be used.
  • a non-displacement type turbine expander as in the sixth embodiment may be used as the expander 5 or the low-stage compressor 1.
  • the combined starting load of the expander 5 and the low-stage compressor 1 can be reduced.
  • the expander 5 or the low-stage compressor 1 is a volumetric expander, for example, a rotary type, scroll type, screw type, vane type, swash plate type, or bankel type.
  • a helical type may be used.
  • a constant flow type gear type, roots type, or screw type may be used as the expander 5 or the low-stage compressor 1.
  • the load at startup is reduced.
  • a positive displacement compressor equipped with a starting load reducing device that can be used may be used to reduce the starting load of the low-stage compressor 1.
  • the starting load reducing device applicable in this case may be a general device such as a method of decelerating the number of rotations at the time of starting, or a method of reducing the suction volume. In this case, if a non-volumetric expander such as a turbine type is used as the expander, the starting load can be further reduced.

Abstract

A freezing cycle apparatus has a two-stage compression device (3) with a low stage side compressor (1) and a high stage side compressor (2), a radiator (4) for cooling a high-pressure gas refrigerant, a pressure reduction device including an expansion device (5) for expanding the high-pressure refrigerant after cooled by the radiator (4), an evaporator (8) for evaporating a low-pressure refrigerant that is reduced in pressure by the pressure reduction device, and a gas injection circuit (10) for introducing an intermediate-pressure gas refrigerant that is reduced in pressure by the pressure reduction device into the suction side of the high stage side compressor (2). The expansion device (5) and the low stage side compressor (1) are coaxially connected.

Description

明 細 書  Specification
冷凍サイクル装置  Refrigeration cycle equipment
技術分野  Technical field
[0001] 本発明は、エネルギー回収用の膨張機を備えた冷凍サイクル装置に関する。  [0001] The present invention relates to a refrigeration cycle apparatus including an expander for energy recovery.
背景技術  Background art
[0002] 近年、冷媒による地球環境破壊の問題力も HFC32、 HFC32を含む混合冷媒など の代替フロンや二酸ィ匕炭素、アンモニアなどの自然冷媒が使用されつつある。ところ 力 Sこのような冷媒を用いた冷凍サイクル装置では高低圧力差が大きくなることから、従 来のフロンを冷媒として用いた冷凍サイクル装置に比し成績係数 (COP)が低 、と ヽ う問題があった。特に、実用化段階を向かえた二酸ィ匕炭素を用いる冷凍サイクル装 置に関して、冷媒を高圧カゝら低圧に減圧する減圧回路中に膨張機を設けることにより エネルギーを回収し、回収したエネルギーを圧縮機の動力に利用して、冷凍サイク ル装置の COPを改善しょうとする試みが多く成されている。  [0002] In recent years, alternative refrigerants such as HFC32 and mixed refrigerants containing HFC32, and natural refrigerants such as carbon dioxide and ammonia are also being used as the problem of global environmental destruction by refrigerants. However, because the refrigeration cycle equipment using such refrigerants has a large difference in pressure between the high and low pressures, the coefficient of performance (COP) is low compared to conventional refrigeration cycle equipment using refrigerants as refrigerants. was there. In particular, with regard to a refrigeration cycle apparatus using carbon dioxide and carbon dioxide that has been put to practical use, energy is recovered by installing an expander in a decompression circuit that decompresses the refrigerant to a low pressure from a high pressure car. Many attempts have been made to improve the COP of refrigeration cycle equipment by using the power of the compressor.
[0003] 特開 2001— 116371号公報はその従来技術において以下のエネルギー回収型 の冷凍サイクル装置を開示する。この装置の基本冷凍サイクルは、一般の冷媒回路 に設けた略等エントロピー膨張を行う容積型膨張機を圧縮機と同軸上に配置し、膨 張機で回収したエネルギーを圧縮機の動力として利用する。ところが、この装置の場 合は、圧縮機を流通する冷媒の質量循環量と膨張機を循環する冷媒の質量循環量 とが等しくなければならず、一方、両者を通過する体積循環量は、それぞれ「シリンダ 容積 X回転数」で決定されることから、次式に示される密度比一定の制約が課せられ る。  [0003] Japanese Patent Laid-Open No. 2001-116371 discloses the following energy recovery type refrigeration cycle apparatus in the prior art. In the basic refrigeration cycle of this device, a positive displacement expander that performs substantially isentropic expansion provided in a general refrigerant circuit is arranged coaxially with the compressor, and the energy recovered by the expander is used as power for the compressor. . However, in the case of this device, the mass circulation amount of the refrigerant flowing through the compressor and the mass circulation amount of the refrigerant circulating through the expander must be equal, while the volume circulation amount passing through both is respectively Since it is determined by “cylinder volume X rotation speed”, the restriction of the constant density ratio shown in the following equation is imposed.
[0004] DEZDC=VCZVE =—定  [0004] DEZDC = VCZVE = —Constant
ここに、 DE:膨張機に吸入される吸入ガス冷媒の密度  Where DE: density of the suction gas refrigerant sucked into the expander
DC：圧縮機に吸入される吸入ガス冷媒の密度  DC: Density of suction gas refrigerant sucked into the compressor
VE：膨張機に吸入される冷媒の体積循環量  VE: Volume circulation volume of refrigerant sucked into expander
VC：圧縮機に吸入される冷媒の体積循環量  VC: Volume circulation volume of refrigerant sucked into the compressor
上式において、 VC及び VEは冷凍サイクル装置固有の値であるので、 VCZVEは 一定値である。 In the above equation, VC and VE are values specific to the refrigeration cycle system, so VCZVE is It is a constant value.
[0005] また、この冷凍サイクル装置を運転した場合は、運転条件により低圧が低下したとき 、モリエル線図上のサイクルの形状が横長四角状力 縦長四角状に変化し、運転効 率が低下する。このため、上記のようにエネルギー回収用膨張機を備えていても効率 の向上を図ることが困難になってくる。  [0005] In addition, when this refrigeration cycle apparatus is operated, when the low pressure is reduced depending on the operating conditions, the shape of the cycle on the Mollier diagram changes to a horizontally long square force and a vertically long square, resulting in a decrease in operating efficiency. . For this reason, it becomes difficult to improve efficiency even if the expander for energy recovery is provided as described above.
[0006] 特開 2001— 116371号公報はこの従来技術における問題を解決する。この発明 は、膨張機と並列に、流量制御弁を介在させたバイパス回路を設けている。そして、 運転条件の変化に対応してバイパス回路を流通する冷媒流量を調節することにより、 膨張機を通過する冷媒流量を調節し、密度比一定と!/ヽぅ冷凍サイクル上の制約を解 消して効率の向上を行っている。  [0006] Japanese Patent Laid-Open No. 2001-116371 solves this problem in the prior art. In the present invention, a bypass circuit having a flow control valve interposed is provided in parallel with the expander. Then, by adjusting the flow rate of refrigerant flowing through the bypass circuit in response to changes in operating conditions, the flow rate of refrigerant passing through the expander is adjusted, and the constant density ratio and the restrictions on the refrigeration cycle are eliminated. To improve efficiency.
発明の開示  Disclosure of the invention
発明が解決しょうとする課題  Problems to be solved by the invention
[0007] この従来技術では、高圧冷媒全てを、膨張機又は絞り機構を通過させて低圧まで 変化させており、このときの冷媒を等エントロピー変化させる膨張機を使用してェネル ギー回収している。し力しながら、実用ィ匕研究が活発に行われている二酸ィ匕炭素を 冷媒に使用した冷凍サイクル装置では、膨張機におけるエネルギーの回収効率は、 あまり大きくないのが実態である。このため、更なる効率改善が要望されている。  [0007] In this prior art, all of the high-pressure refrigerant is changed to low pressure through an expander or a throttle mechanism, and energy is recovered using an expander that changes the refrigerant at this time isentropically. . However, in the refrigeration cycle equipment using carbon dioxide as a refrigerant, which is actively researched for practical use, the actual energy recovery efficiency of the expander is not so high. For this reason, further efficiency improvement is desired.
[0008] 本発明は、膨張機によるエネルギー回収を行うとともに、圧縮機における消費エネ ルギ一の低減を併せ行うことにより、効率向上を行った冷凍サイクル装置を提供する ことを目的とする。  [0008] It is an object of the present invention to provide a refrigeration cycle apparatus that improves efficiency by performing energy recovery using an expander and reducing energy consumption in the compressor.
課題を解決するための手段  Means for solving the problem
[0009] 本発明に係る冷凍サイクル装置は上記課題を解決するためになされたものであつ て、第 1の発明は、低段側圧縮機と高段側圧縮機とからなる 2段圧縮装置と、高圧ガ ス冷媒を冷却する放熱器と、放熱器で冷却された後の高圧冷媒を膨張させる膨張機 を含む減圧装置と、減圧装置により減圧された低圧冷媒を蒸発させる蒸発器と、減 圧装置で減圧された中間圧ガス冷媒を高段側圧縮の吸入側に導入するガスインジ ェクシヨン回路とを備え、前記膨張機と低段側圧縮機とは同軸で連結されていること を特徴とする。 [0010] 中間圧ガス冷媒を高段側圧縮機の吸入側に導入するガスインジェクション回路を 設けているので、従来技術 1及び 2に比較し高圧ガス冷媒を減圧する膨張機におい て回収するエネルギーが減少するが、 2段圧縮装置における圧縮仕事量が減少する 。また、従来技術におけるエネルギー回収効率はあまり大きくない。このような要因に より、本発明は、圧縮装置におけるエネルギー消費量を低減させることにより、ェネル ギー効率を向上させることができる。 [0009] The refrigeration cycle apparatus according to the present invention has been made to solve the above-mentioned problems, and the first invention is a two-stage compression apparatus comprising a low-stage compressor and a high-stage compressor. A decompressor including a radiator that cools the high-pressure gas refrigerant, an expander that expands the high-pressure refrigerant after being cooled by the radiator, an evaporator that evaporates the low-pressure refrigerant decompressed by the decompressor, and a decompression A gas injection circuit that introduces the intermediate-pressure gas refrigerant decompressed by the apparatus to the suction side of the high-stage compression, and the expander and the low-stage compressor are connected coaxially. [0010] Since the gas injection circuit for introducing the intermediate-pressure gas refrigerant to the suction side of the high-stage compressor is provided, the energy recovered in the expander that depressurizes the high-pressure gas refrigerant as compared with the prior arts 1 and 2. Although it decreases, the compression work in the two-stage compressor decreases. Further, the energy recovery efficiency in the prior art is not so great. Due to such factors, the present invention can improve the energy efficiency by reducing the energy consumption in the compressor.
図面の簡単な説明  Brief Description of Drawings
[0011] [図 1]実施の形態 1に係る冷凍サイクル装置の冷媒回路図である。 FIG. 1 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 1.
[図 2]同冷凍サイクル装置のモリエル線図である。  FIG. 2 is a Mollier diagram of the refrigeration cycle apparatus.
[図 3]実施の形態 1に係る冷凍サイクル装置における冷凍 (又は冷房) COP比線図で ある。  FIG. 3 is a refrigeration (or cooling) COP ratio diagram in the refrigeration cycle apparatus according to Embodiment 1.
[図 4]実施の形態 2に係る冷凍サイクル装置の冷媒回路図である。  FIG. 4 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 2.
[図 5]実施の形態 3に係る冷凍サイクル装置の冷媒回路図である。  FIG. 5 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 3.
[図 6]実施の形態 4に係る冷凍サイクル装置の冷媒回路図である。  FIG. 6 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 4.
[図 7]同冷凍サイクル装置のモリエル線図である。  FIG. 7 is a Mollier diagram of the refrigeration cycle apparatus.
[図 8]実施の形態 5に係る冷凍サイクル装置における給湯 (又は暖房) COP比線図で ある。  FIG. 8 is a hot water supply (or heating) COP ratio diagram in the refrigeration cycle apparatus according to Embodiment 5.
[図 9]実施の形態 6に係る冷凍サイクル装置の冷媒回路図である。  FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiment 6.
発明を実施するための最良の形態  BEST MODE FOR CARRYING OUT THE INVENTION
[0012] (第 1の実施の形態） [0012] (First embodiment)
以下に、本発明の第 1の実施の形態を図 1に基づいて説明する。冷凍サイクル装置 は、低段側圧縮機 1と高段側圧縮機 2とからなる 2段圧縮装置 3を備える。同圧縮装 置 3から放出される高圧ガス冷媒は放熱器 4にて冷却されたのち、膨張機 5にて減圧 された中間圧冷媒として気液分離器 6にて気液分離される。気液分離器 6にてガスと 分離された液体は絞り装置 7にて低圧に減圧されたのち蒸発器 8にて蒸発気化され てアキュムレータ 9を経て低段側圧縮機 1の吸入ポートに送られる。一方、気液分離 器 6において冷媒液と分離されたガスはガスインジェクション回路 10を経て高段側圧 縮機 2の吸入ポートに送られる。 [0013] 上記冷凍サイクル装置において、低段側及び高段側の圧縮機 1, 2はロータリ圧縮 機のような容量一定の圧縮機である。そして、低段側圧縮機 1と膨張機 5とは同軸に 連結されている。また、ガスインジヱクシヨン回路 10の開閉弁 12の開度を調節するこ とにより、同回路 10の流量を適切に設定するようにしてもよい。また、バイパス回路 11 の流量制御弁 13は低段側圧縮機 1の吸入圧力により開度が制御される。即ち、低段 側圧縮機 1に吸入される冷媒の圧力が上昇すれば流量制御弁 13の開度が大きなる ように制御されている。なお、冷凍サイクル装置では、膨張機 5と絞り装置 7とが本発 明における減圧装置を構成する。また、このように構成された冷凍サイクル装置には 、冷媒として HFC32又は HFC32などの通常の冷凍サイクルを形成する冷媒 (超臨 界冷凍サイクルを形成しな ヽ冷媒）が充填されて 、る。 Hereinafter, a first embodiment of the present invention will be described with reference to FIG. The refrigeration cycle apparatus includes a two-stage compressor 3 composed of a low-stage compressor 1 and a high-stage compressor 2. The high-pressure gas refrigerant released from the compression device 3 is cooled by the radiator 4 and then separated by the gas-liquid separator 6 as an intermediate-pressure refrigerant decompressed by the expander 5. The liquid separated from the gas by the gas-liquid separator 6 is depressurized to a low pressure by the expansion device 7 and then evaporated and evaporated by the evaporator 8 and sent to the suction port of the low-stage compressor 1 via the accumulator 9. . On the other hand, the gas separated from the refrigerant liquid in the gas-liquid separator 6 is sent to the suction port of the high-stage compressor 2 through the gas injection circuit 10. [0013] In the refrigeration cycle apparatus, the low-stage and high-stage compressors 1 and 2 are constant capacity compressors such as a rotary compressor. The low stage compressor 1 and the expander 5 are connected coaxially. Further, the flow rate of the circuit 10 may be appropriately set by adjusting the opening degree of the on-off valve 12 of the gas instruction circuit 10. Further, the opening degree of the flow control valve 13 of the bypass circuit 11 is controlled by the suction pressure of the low-stage compressor 1. That is, the flow control valve 13 is controlled so that the opening degree of the refrigerant is increased when the pressure of the refrigerant sucked into the low-stage compressor 1 increases. In the refrigeration cycle apparatus, the expander 5 and the expansion device 7 constitute a decompression device in the present invention. The refrigeration cycle apparatus configured in this manner is filled with a refrigerant that forms a normal refrigeration cycle such as HFC32 or HFC32 (a refrigerant that does not form a supercritical refrigeration cycle) as a refrigerant.
[0014] 冷凍サイクル装置の動作を、図 2に従って説明する。低段側圧縮機 1から吐出され た吐出ガス (al)は、気液分離器 6からガスインジェクション回路 10を経て高段側圧 縮機 2の吸入側に導入された中間圧ガス冷媒 (a8)と混合して (a2)、高段側圧縮機 2 に吸入される。高段側圧縮機 2から吐出された高圧ガス冷媒 (a3)は放熱器 4で冷却 される (a4) G放熱器 4で冷却された高圧冷媒は膨張機 5により略等ェンタルピー変化 の減圧作用を受けて中間圧まで減圧される (a5)。このとき、膨張機 5と低段側圧縮機 1とが同軸に接続されていることにより、膨張機 5により回収されたエネルギーが低段 側圧縮機 1を回転するのに消費されるエネルギーとして利用される。 [0014] The operation of the refrigeration cycle apparatus will be described with reference to FIG. The discharge gas (al) discharged from the low-stage compressor 1 and the intermediate-pressure gas refrigerant (a8) introduced from the gas-liquid separator 6 through the gas injection circuit 10 to the suction side of the high-stage compressor 2 Mix (a2) and suck into high-stage compressor 2. The high-pressure gas refrigerant (a3) discharged from the high-stage compressor 2 is cooled by the radiator 4 (a4) The high-pressure refrigerant cooled by the G radiator 4 is decompressed by the expander 5 and has a substantially isenthalpy change. In response, the pressure is reduced to an intermediate pressure (a5). At this time, since the expander 5 and the low-stage compressor 1 are connected coaxially, the energy recovered by the expander 5 is used as energy consumed to rotate the low-stage compressor 1. Is done.
[0015] 中間圧まで減圧された冷媒 (a5)は気液分離器 6で気液分離される。分離されたガ ス冷媒 (a8)は前述のように高段側圧縮機 2の吸入側に導入されて低段側圧縮機 1 の吐出ガス (al)と混合される（a2)。一方、気液分離器 6で分離された液冷媒 (a7)は 、絞り装置 7により減圧されて低圧冷媒 (a8)となる。この低圧冷媒は蒸発器 8で蒸発 気化され (a9)、アキュムレータ 9を経て低段側圧縮機 1に吸入される。このアキュムレ ータ 9は、低段側圧縮機 1の吸入側に未蒸発の液冷媒が流れてきた場合に、この液 冷媒を分離して貯留することにより、低段側圧縮機 1における液圧縮を防止するもの である。  The refrigerant (a5) decompressed to the intermediate pressure is gas-liquid separated by the gas-liquid separator 6. The separated gas refrigerant (a8) is introduced to the suction side of the high-stage compressor 2 as described above and mixed with the discharge gas (al) of the low-stage compressor 1 (a2). On the other hand, the liquid refrigerant (a7) separated by the gas-liquid separator 6 is decompressed by the expansion device 7 to become a low-pressure refrigerant (a8). This low-pressure refrigerant is evaporated and evaporated by the evaporator 8 (a9), and is sucked into the low-stage compressor 1 through the accumulator 9. The accumulator 9 separates and stores the liquid refrigerant that has not evaporated in the suction side of the low-stage compressor 1 to store the liquid refrigerant in the low-stage compressor 1. This is to prevent this.
[0016] また、冷凍サイクル装置にぉ 、ては、蒸発温度を運転許容範囲における最低温度（ 例えば— 35°C)とした場合に、バイパス回路 11の流量制御弁 13の開度を最少（つま り閉鎖状態)としておくことで、密度比一定の制約を満たすようにする。そして、蒸発 器 8の蒸発負荷が増大して低段側圧縮機 1の吸入圧力の増加とともに高段側圧縮機 2の吸入圧力が上昇した場合に、その増加に対応して流量制御弁 13の開度を大きく してバイパス回路 11の冷媒循環量を増大させるように制御する。これにより、放熱器 4で冷却された高圧冷媒のうち膨張機 5を通過させない余剰冷媒を、バイパス回路 1 1を通じて気液分離器 6に流す。このとき、この高圧冷媒は流量制御弁 13により中間 圧まで減圧される (a6)。そして、前述の膨張機 5を介して気液分離器 6に流されてき た中間圧冷媒と混合し、気液分離器 6で気液分離される。 [0016] Further, in the refrigeration cycle apparatus, when the evaporation temperature is set to the lowest temperature within the allowable operating range (for example, -35 ° C), the opening degree of the flow control valve 13 of the bypass circuit 11 is minimized (that is, (Closed state) to satisfy the constraint of a constant density ratio. When the evaporation load of the evaporator 8 increases and the suction pressure of the high-stage compressor 2 increases as the suction pressure of the low-stage compressor 1 increases, the flow control valve 13 corresponds to the increase. Control is performed to increase the amount of refrigerant circulation in the bypass circuit 11 by increasing the opening. Thus, surplus refrigerant that does not pass through the expander 5 among the high-pressure refrigerant cooled by the radiator 4 is caused to flow to the gas-liquid separator 6 through the bypass circuit 11. At this time, the high-pressure refrigerant is reduced to an intermediate pressure by the flow control valve 13 (a6). Then, it is mixed with the intermediate-pressure refrigerant that has flowed to the gas-liquid separator 6 via the expander 5 described above, and the gas-liquid separator 6 performs gas-liquid separation.
[0017] 上記構成によれば、膨張機 5と並列に流量制御弁 13を備えたノ ィパス回路 11を設 け、蒸発温度の上昇に対応してバイパス量を増加するように制御することにより密度 比一定の制約力も解放されるので、モリエル線図が非効率的な縦長の線図になるこ とを防止することができ、効率を向上することができる。  [0017] According to the above configuration, the density is controlled by providing the no-pass circuit 11 including the flow control valve 13 in parallel with the expander 5 and controlling the amount of bypass to increase in response to the increase in the evaporation temperature. Since the constant constraint force is also released, it is possible to prevent the Mollier diagram from becoming an inefficient vertical diagram, and to improve efficiency.
[0018] また、この冷凍サイクル装置によれば、中間圧ガス冷媒を高段側圧縮機 2の吸入側 に導入するガスインジェクション回路 10を設けているので、従来技術に比較し高圧ガ ス冷媒を減圧する膨張機 5における回収エネルギーが減少するが、 2段圧縮装置 3 における圧縮仕事量が従来技術よりも減少する。また、膨張機 5によるエネルギー回 収効率が低いことから、膨張機 5による回収エネルギーの減少よりも 2段圧縮装置 3に おける圧縮エネルギーの減少量のほうが大きくなる結果、従来技術より効率が改善さ れる。図 3はこの結果を示した例である。  [0018] Further, according to this refrigeration cycle apparatus, since the gas injection circuit 10 for introducing the intermediate-pressure gas refrigerant to the suction side of the high-stage compressor 2 is provided, the high-pressure gas refrigerant is used as compared with the prior art. Although the recovered energy in the decompressing expander 5 is reduced, the compression work in the two-stage compressor 3 is reduced as compared with the prior art. In addition, since the energy recovery efficiency by the expander 5 is low, the amount of decrease in the compression energy in the two-stage compressor 3 is larger than the decrease in the recovered energy by the expander 5, resulting in improved efficiency over the prior art. It is. Figure 3 shows an example of this result.
[0019] 図 3は、実施の形態 1の冷凍サイクル装置を冷凍装置 (又は冷房装置）とした場合 において、蒸発温度の変化に対応する冷凍 (又は冷房) COP比を、従来技術と比較 できるように示している。この場合における冷凍 (又は冷房) COP比とは、膨張機 5に よるエネルギー回収を行わない古くからある一般的な冷凍装置 (又は冷房装置)の各 蒸発温度における冷凍 (又は冷房) COPに対する、本願の第 1及び第 2の実施の形 態における COPの比で表したものである。この図に示すように、蒸発温度が低いとき は高段側圧縮機 2の吸入圧力も低くなり、バイパス回路 11からバイパスされる冷媒量 が減少するので、膨張機 5によるエネルギー回収量が多くなり COP比が大きくなる。 また、従来技術と比較すると、膨張機 5による回収エネルギーの減少よりも 2段圧縮装 置 3における圧縮エネルギーの減少量が大きくなる結果、従来技術の COP比に対し て COP比が大きくなつて!/、る。 FIG. 3 shows that when the refrigeration cycle apparatus of Embodiment 1 is a refrigeration apparatus (or a cooling apparatus), the refrigeration (or cooling) COP ratio corresponding to the change in the evaporation temperature can be compared with the prior art. It shows. The refrigeration (or cooling) COP ratio in this case refers to the refrigeration (or cooling) COP at each evaporating temperature of an old general refrigeration apparatus (or cooling apparatus) that does not perform energy recovery by the expander 5. This is expressed as the ratio of COP in the first and second embodiments. As shown in this figure, when the evaporation temperature is low, the suction pressure of the high-stage compressor 2 is also reduced, and the amount of refrigerant bypassed from the bypass circuit 11 is reduced, so that the amount of energy recovered by the expander 5 is increased. COP ratio increases. In addition, compared with the conventional technology, the two-stage compressor As a result, the amount of decrease in compression energy in device 3 increases, resulting in a larger COP ratio than the conventional COP ratio!
[0020] (第 2の実施の形態）  [0020] (Second Embodiment)
次に、第 2の実施の形態について、図 4に従って説明する。第 2の実施の形態は、 第 1の実施の形態の低段側圧縮機 1を容量可変型として、流量制御弁 13を備えたバ ィパス回路 11、及びガスインジェクション回路 10中の開閉弁 12を省略したものであ る。また、低段側圧縮機 1と膨張機 5とは同一ケーシング 21内に収納されている。即 ち、ケーシング 21内を隔壁 22により 2室に区画し、一方の室 23には低段側圧縮機 1 が収納され、他方の部屋 24には膨張機 5が収納されている。アキュムレータ 9の出口 ポートは圧縮機室 23内に接続され、低段側圧縮機 1はこの圧縮機室 23内の冷媒を 吸入する。また、この冷凍サイクル装置におけるモリエル線図は、図 2と一部で異なる 。すなわち、この実施の形態ではバイノス回路 11が省略されているので、モリエル線 図も、図 2において、流量制御弁 13による略等ェンタルピー変化（a4→a6)を示す破 線部分を省略した線図となる。なお、前述の容量可変型の低段側圧縮機 1として容 量可変型の斜板式圧縮機を用いて!/ヽる。  Next, a second embodiment will be described with reference to FIG. In the second embodiment, the low-stage compressor 1 of the first embodiment is of a variable capacity type, and a bypass circuit 11 having a flow control valve 13 and an on-off valve 12 in the gas injection circuit 10 are provided. It is omitted. Further, the low-stage compressor 1 and the expander 5 are accommodated in the same casing 21. That is, the casing 21 is divided into two chambers by the partition wall 22, the low-stage compressor 1 is stored in one chamber 23, and the expander 5 is stored in the other chamber 24. The outlet port of the accumulator 9 is connected to the compressor chamber 23, and the low-stage compressor 1 sucks the refrigerant in the compressor chamber 23. In addition, the Mollier diagram in this refrigeration cycle apparatus is partially different from FIG. In other words, since the binos circuit 11 is omitted in this embodiment, the Mollier diagram is also a diagram in which the broken line portion showing the substantially isenthalpy change (a4 → a6) by the flow control valve 13 is omitted in FIG. It becomes. In addition, a variable capacity swash plate compressor is used as the variable capacity low stage compressor 1 described above.
[0021] また、容量可変型の低段側圧縮機 1は次のように制御される。すなわち、蒸発温度 を運転許容範囲における最低温度 (例えば— 35°C)とした場合に、低段側圧縮機 1 の回転数が最大となるように設定されている。そして、蒸発器 8の蒸発負荷が増大す ると、低段側圧縮機 1の吸入圧力が上昇する。これに伴う高段側圧縮機 2の吸入圧 力の上昇に対応して低段側圧縮機 1の容量が減少するように制御される。  [0021] The variable-capacity low-stage compressor 1 is controlled as follows. That is, when the evaporation temperature is the lowest temperature in the allowable operating range (for example, −35 ° C.), the rotation speed of the low-stage compressor 1 is set to be the maximum. When the evaporation load of the evaporator 8 increases, the suction pressure of the low-stage compressor 1 increases. The capacity of the low-stage compressor 1 is controlled to decrease in response to the increase in the suction pressure of the high-stage compressor 2 accompanying this.
[0022] 第 2の実施の形態 2によれば、低段側圧縮機 1を容量可変型圧縮機として、運転条 件の変化に対応して低段側圧縮機 1の能力を制御することにより、密度比一定の制 約から解放されるので、モリエル線図が非効率的な縦長の線図になることを防止する ことができる。また、第 1の実施の形態 1の場合のように膨張機をバイパスして密度比 一定の制約カゝら解放されるものに較べて、低段側圧縮機 1の容量を制御することで、 より一層の効率改善を図ることができる。  [0022] According to the second embodiment, the low-stage compressor 1 is used as a variable displacement compressor, and the capacity of the low-stage compressor 1 is controlled in response to changes in operating conditions. In addition, since the density ratio is fixed, the Mollier diagram can be prevented from becoming an inefficient vertical diagram. In addition, by controlling the capacity of the low-stage compressor 1 as compared to the case where the expansion unit is bypassed and the constant density ratio is released as in the case of the first embodiment 1, The efficiency can be further improved.
[0023] また、低段側圧縮機 1が容量可変型斜板式圧縮機とされているので、膨張機 5に流 入する冷媒の密度比と高段側圧縮機 2に流入する冷媒の密度比とを無段階に調節 することが可能となり、より一層エネルギー効率を向上させることができる。 [0023] Since the low-stage compressor 1 is a variable capacity swash plate compressor, the density ratio of the refrigerant flowing into the expander 5 and the density ratio of the refrigerant flowing into the high-stage compressor 2 And adjust steplessly Energy efficiency can be further improved.
[0024] (第 3の実施の形態）  [0024] (Third embodiment)
次に、第 3の実施の形態について、図 5に従って説明する。第 3の実施の形態にお いては、第 2の実施の形態の低段側圧縮機 1と、気液分離器 6と、アキュムレータ 9と を一体ィ匕したものである。すなわち、第 3の実施の形態においては、圧縮機室 23の 下部をアキュムレータとしている。このために、蒸発器 8の出口ポートを直接に圧縮機 室 23に対して接続するとともに、低段側圧縮機 1は圧縮機室 23内の冷媒を直接に 吸入する。従って、蒸発器 8から帰還する液冷媒は圧縮機室 23の下部に貯留される ことにより、圧縮機室 23がアキュムレータとしても機能する。  Next, a third embodiment will be described with reference to FIG. In the third embodiment, the low-stage compressor 1, the gas-liquid separator 6, and the accumulator 9 of the second embodiment are integrated. That is, in the third embodiment, the lower part of the compressor chamber 23 is an accumulator. For this purpose, the outlet port of the evaporator 8 is directly connected to the compressor chamber 23, and the low-stage compressor 1 directly sucks the refrigerant in the compressor chamber 23. Therefore, the liquid refrigerant returning from the evaporator 8 is stored in the lower part of the compressor chamber 23, so that the compressor chamber 23 also functions as an accumulator.
[0025] また、この実施の形態においては、膨張機室 24が気液分離器としても使用されて いる。このために、膨張機 5の出口ポートを膨張機室 24内に開放するとともに、膨張 機室 24の一部が蒸発器 8に絞り装置 7を介し接続されている。この結果、膨張機 5を 通過した中間圧冷媒が圧縮機室 23内に放出される。そして、圧縮機室 23の空間内 で重力差により気液混合の中間圧冷媒が気液分離される。また、分離された液冷媒 は、絞り装置 7を介して蒸発器 8に流出される。なお、この実施の形態のモリエル線図 は、実施の形態 2の場合と同様であり、蒸発温度の変化に対するエネルギー効率は 図 3に示される。  [0025] In this embodiment, the expander chamber 24 is also used as a gas-liquid separator. For this purpose, the outlet port of the expander 5 is opened in the expander chamber 24, and a part of the expander chamber 24 is connected to the evaporator 8 via the expansion device 7. As a result, the intermediate pressure refrigerant that has passed through the expander 5 is discharged into the compressor chamber 23. Then, the intermediate-pressure refrigerant in the gas-liquid mixture is separated into gas and liquid by the difference in gravity in the space of the compressor chamber 23. Further, the separated liquid refrigerant flows out to the evaporator 8 through the expansion device 7. The Mollier diagram of this embodiment is the same as that of Embodiment 2, and the energy efficiency with respect to the change in evaporation temperature is shown in FIG.
[0026] 第 3の実施の形態においては、段側圧縮機 1と膨張機 5とが隔壁 22にて区画された 同一のケーシング 21内に収納され、さらに、低段側圧縮機 1の収納された圧縮機室 23内がアキュムレータとして兼用されている。従って、第 3の実施の形態における冷 凍サイクル装置では、低段側圧縮機 1、膨張機 5及びアキュムレータを個別に製作し て取り付ける場合に比し、機器がコンパクトになるとともに、アキュムレータを他の機器 に接続する配管が簡素化される。  In the third embodiment, the stage side compressor 1 and the expander 5 are accommodated in the same casing 21 partitioned by the partition wall 22, and further, the low stage side compressor 1 is accommodated. The compressor chamber 23 is also used as an accumulator. Therefore, in the refrigeration cycle apparatus according to the third embodiment, the apparatus is more compact and the accumulator can be installed in other units as compared to the case where the low-stage compressor 1, the expander 5 and the accumulator are individually manufactured and installed. Piping connected to equipment is simplified.
[0027] さらに、前記膨張機 5を収納した膨張機室 24が気液分離器として使用されるので、 低段側圧縮機、膨張機、気液分離器及びアキュムレータを個別に製作して取り付け る冷凍サイクル装置と比し、機器がコンパクトになるとともに、これら機器を接続する配 管が簡素化される。  [0027] Further, since the expander chamber 24 containing the expander 5 is used as a gas-liquid separator, a low-stage compressor, an expander, a gas-liquid separator, and an accumulator are individually manufactured and attached. Compared to the refrigeration cycle device, the equipment is compact and the piping connecting these equipment is simplified.
[0028] (第 4の実施の形態） 次に、第 4の実施の形態に係る冷凍サイクル装置について、図 6, 7に従って説明 する。冷凍サイクル装置は、低段側圧縮機 1と高段側圧縮機 2とからなる 2段圧縮装 置 3、放熱器 4、絞り装置 31、絞り装置 31を通過した中間圧冷媒と残部の高圧冷媒と を熱交換させる熱交換器 32、熱交換器 32で冷却された中間圧冷媒を低圧まで減圧 する膨張機 33、蒸発器 8、アキュムレータ 9などカゝら構成されている。なお、熱交 32は、絞り装置 31を通過した中間圧冷媒と残部の高圧冷媒とを熱交換させるために 、高圧冷媒を流通させる高圧側経路 32aと中間圧冷媒を流通させる中間圧側経路 3 2bとが近接して配置されている。熱交翻32の高圧側経路 32aの入口は放熱器 4 の出口に接続され、高圧側経路 32aの出口は膨張機 33に接続されている。また、中 間圧側経路 32bの入口は絞り装置 31に、出口はガスインジェクション回路 34により 高段側圧縮機 2の吸入ポートに対してそれぞれ接続されている。 [0028] (Fourth embodiment) Next, a refrigeration cycle apparatus according to a fourth embodiment will be described with reference to FIGS. The refrigeration cycle equipment consists of a two-stage compression device 3 consisting of a low-stage compressor 1 and a high-stage compressor 2, a radiator 4, a throttling device 31, a throttling device 31, an intermediate-pressure refrigerant and the remaining high-pressure refrigerant. A heat exchanger 32 for exchanging heat with each other, an expander 33 for reducing the intermediate pressure refrigerant cooled by the heat exchanger 32 to a low pressure, an evaporator 8 and an accumulator 9 are configured. In order to exchange heat between the intermediate-pressure refrigerant that has passed through the expansion device 31 and the remaining high-pressure refrigerant, the heat exchange 32 is connected to the high-pressure side path 32a through which the high-pressure refrigerant flows and the intermediate pressure-side path 32b through which the intermediate pressure refrigerant flows. And are arranged close to each other. The inlet of the high-pressure side path 32 a of the heat exchanger 32 is connected to the outlet of the radiator 4, and the outlet of the high-pressure side path 32 a is connected to the expander 33. Further, the inlet of the intermediate pressure side passage 32b is connected to the expansion device 31, and the outlet is connected to the suction port of the high stage compressor 2 by the gas injection circuit 34.
[0029] また、低段側圧縮機 1は容量可変型の斜板式圧縮機であり、高段側圧縮機 2は口 一タリ圧縮機のような容量一定の圧縮機である。低段側圧縮機 1と膨張機 33とは同 軸に連結されて 、る。冷媒として HFC32又は HFC32などの通常の冷凍サイクルを 形成する冷媒 (超臨界冷凍サイクルを形成しな!ヽ冷媒)が充填されて!ヽるものとする。  [0029] The low-stage compressor 1 is a variable capacity swash plate compressor, and the high-stage compressor 2 is a constant capacity compressor such as a single compressor. The low-stage compressor 1 and the expander 33 are connected to the same shaft. It shall be filled with a refrigerant that forms a normal refrigeration cycle such as HFC32 or HFC32 (a refrigerant that does not form a supercritical refrigeration cycle).
[0030] 上記した冷凍サイクル装置の動作を、図 6に従って説明する。低段側圧縮機 1から 吐出された吐出ガス (bl)は、熱交 32からの中間圧ガス冷媒 (b9)と混合して (b [0030] The operation of the above-described refrigeration cycle apparatus will be described with reference to FIG. The discharge gas (bl) discharged from the low-stage compressor 1 is mixed with the intermediate-pressure gas refrigerant (b9) from the heat exchanger 32 (b
2)、高段側圧縮機 2に吸入される。高段側圧縮機 2から吐出された高圧ガス冷媒 (b2), sucked into high stage compressor 2 High-pressure gas refrigerant discharged from high-stage compressor 2 (b
3)は放熱器 4で冷却される (b4)。放熱器 4で冷却された一部の高圧冷媒 (b4)は絞 り装置 31により略等ェンタルピー変化の減圧作用を受けて中間圧まで減圧される (b3) is cooled by radiator 4 (b4). A part of the high-pressure refrigerant (b4) cooled by the radiator 4 is decompressed to an intermediate pressure by the throttling device 31 under a pressure-reducing action of a substantially isenthalpy change (b
8)。減圧された冷媒 (b8)は、熱交 で残部の高圧冷媒と熱交換して気化し (b8). The decompressed refrigerant (b8) is vaporized by exchanging heat with the remaining high-pressure refrigerant by heat exchange (b
9)、低段側圧縮機 1からの吐出ガス (bl)と混合して (b2)、高段側圧縮機 2の吸入側 に導入される。 9) Mixed with the discharge gas (bl) from the low stage compressor 1 (b2) and introduced into the suction side of the high stage compressor 2.
[0031] 一方、熱交換器 32で冷却された高圧冷媒 (b5)は、膨張機 33により略等エントロピ 一変化を受けて低圧まで減圧される (b6)。このとき、膨張機 33と低段側圧縮機 1とが 同軸に接続されていることにより、膨張機 33により回収されたエネルギーが低段側圧 縮機 1を回転するためのエネルギーとして利用される。低圧冷媒 (b6)は、蒸発器 8で 蒸発気化して (b7)、低段側圧縮機 1に吸入される。 [0032] また、容量可変型の低段側圧縮機 1は、第 2の実施の形態と同様に制御される。す なわち、蒸発温度を運転許容範囲における最低温度 (例えば— 35°C)とした場合に 、低段側圧縮機 1の回転数が最大となるように設定することで、密度比一定の制約を 満たすようにする。蒸発器 8の蒸発負荷が増大して高段側圧縮機 2の吸入圧力が上 昇した場合に、その増加に対応して低段側圧縮機 1の容量が減少される。 On the other hand, the high-pressure refrigerant (b5) cooled by the heat exchanger 32 is decompressed to a low pressure by undergoing a substantially isentropic change by the expander 33 (b6). At this time, since the expander 33 and the low-stage compressor 1 are connected coaxially, the energy recovered by the expander 33 is used as energy for rotating the low-stage compressor 1. The low-pressure refrigerant (b6) evaporates and evaporates in the evaporator 8 (b7) and is sucked into the low-stage compressor 1. [0032] The variable capacity low stage compressor 1 is controlled in the same manner as in the second embodiment. In other words, when the evaporation temperature is set to the lowest temperature within the allowable operating range (for example, -35 ° C), by setting so that the rotation speed of the low-stage compressor 1 is maximized, the constant density ratio is restricted. To satisfy. When the evaporation load of the evaporator 8 increases and the suction pressure of the high-stage compressor 2 increases, the capacity of the low-stage compressor 1 is reduced corresponding to the increase.
[0033] 上記の冷凍サイクル装置では、低段側圧縮機 1が容量可変型斜板式圧縮機とされ ているので、運転条件の変化に対応して圧縮機の吸入側の密度比を調節することが 可能となり、モリエル線図が非効率的な縦長の線図になることが防止される。  [0033] In the above refrigeration cycle apparatus, the low-stage compressor 1 is a variable displacement swash plate compressor, and therefore, the density ratio on the suction side of the compressor is adjusted in response to changes in operating conditions. This prevents the Mollier diagram from becoming an inefficient vertical diagram.
[0034] また、放熱器 4で冷却された一部の高圧冷媒を絞り装置 31で中間圧まで減圧し、こ の減圧した気液混合の冷媒と、残部の高圧冷媒とを熱交換器 32で熱交換させて ヽ る。よって、膨張機 33を経由して蒸発器 8に流通させる冷媒のェンタルピーを小さく することができ、冷凍効率が改善される。  [0034] Further, a part of the high-pressure refrigerant cooled by the radiator 4 is decompressed to an intermediate pressure by the expansion device 31, and the decompressed gas-liquid mixed refrigerant and the remaining high-pressure refrigerant are separated by the heat exchanger 32. Let the heat exchange. Therefore, the enthalpy of the refrigerant flowing through the expander 33 to the evaporator 8 can be reduced, and the refrigeration efficiency is improved.
[0035] また、熱交翻 32の高圧側経路と蒸発器 8との間に、放熱器 4及び熱交翻 32で 冷却された後の残部の高圧冷媒を低圧になるまで膨張させる膨張機 33を設けてい る。従って、冷媒の減圧に相当するエネルギーが膨張機 33で回収される。この場合 、膨張機 33で回収されるエネルギーは中間圧ガス冷媒として、ガスインジェクション 回路 34を通じて高段側圧縮機 2に流通される。従って、熱交換器 32を介して高段側 圧縮機 2に導入される冷媒循環量分だけ低段側圧縮機 1における圧縮仕事量が減 少するため、エネルギー効率が改善される。  [0035] Further, the expander 33 expands the remaining high-pressure refrigerant after being cooled by the radiator 4 and the heat exchanger 32 to a low pressure between the high-pressure side path of the heat exchanger 32 and the evaporator 8. Is provided. Accordingly, energy corresponding to the decompression of the refrigerant is recovered by the expander 33. In this case, the energy recovered by the expander 33 is circulated to the high-stage compressor 2 through the gas injection circuit 34 as an intermediate pressure gas refrigerant. Accordingly, the amount of compression work in the low-stage compressor 1 is reduced by the amount of refrigerant circulating introduced into the high-stage compressor 2 via the heat exchanger 32, so that energy efficiency is improved.
[0036] なお、この実施の形態 4については、図 3において、蒸発温度の変化に対する CO P比の変化を特に示していないが、実施の形態 2及び 3と同等の COP比を得ており、 従来技術及び第 1の実施の形態に比し高効率の運転を行うことができる。  [0036] Note that in this Embodiment 4, the change in COP ratio with respect to the change in evaporation temperature is not particularly shown in Fig. 3, but the COP ratio equivalent to that in Embodiments 2 and 3 was obtained. Compared with the prior art and the first embodiment, a highly efficient operation can be performed.
[0037] (第 5の実施の形態）  [0037] (Fifth embodiment)
次に第 5の実施の形態について説明する。この実施の形態は、第 1の実施の形態 の冷媒回路において、冷媒として二酸ィ匕炭素が使用される冷凍サイクル装置である 。なお、放熱器 4において水が加熱され、これで得た温水が給湯又は暖房に使用さ れる。なお、この場合のモリエル線図は、図 2のようなサイクルではある力 超臨界サ イタルとなり、高段側圧縮機 2出口の吐出ガス及び放熱器 4で冷却された高圧冷媒は 、何れも臨界点以上の圧力の高圧ガス冷媒となる。したがって、高圧側回路には液 冷媒の存在する余地がないので、アキュムレータ 9で冷媒量が調節される。このため に、アキュムレータ 9の大きさを、余剰冷媒を貯留可能とする大きさに設定している。 Next, a fifth embodiment will be described. This embodiment is a refrigeration cycle apparatus using carbon dioxide as a refrigerant in the refrigerant circuit of the first embodiment. In addition, water is heated in the radiator 4 and the hot water obtained by this is used for hot water supply or heating. The Mollier diagram in this case is a force supercritical cycle in the cycle as shown in Fig. 2, and the high-pressure refrigerant discharged from the high-stage compressor 2 outlet and the high-pressure refrigerant cooled by the radiator 4 are These are high-pressure gas refrigerants having a pressure higher than the critical point. Therefore, there is no room for liquid refrigerant in the high-pressure circuit, and the amount of refrigerant is adjusted by the accumulator 9. For this purpose, the size of the accumulator 9 is set to a size that allows excess refrigerant to be stored.
[0038] したがって、この冷凍サイクル装置によれば、放熱器 4で加熱された温水を給湯又 は暖房に使用することにより、高温の給湯水、又は高温の吹き出し空気を得ることが できる。また、二酸ィ匕炭素を冷媒として使用した冷凍サイクル装置としては、従来に比 し大幅な効率の改善を行うことができる。図 8は、実施の形態 5の冷凍サイクル装置に おいて、放熱器 4出口の温水温度の変化（間接的には、高圧の変化）に対する給湯（ 又は暖房) COP比の効果を示して 、る。この場合における給湯 (又は暖房) COP比と は、膨張機 5によるエネルギー回収を行わない古くからある一般的な冷凍サイクル装 置を応用した給湯装置 (又は暖房装置)の凝縮器出口の温水温度が変化した場合 の給湯 (又は暖房) COPに対する、実施の形態 5の同 COPの比を表したものである。 図示するように、放熱器 4出口の温水温度が低いときは（つまり高圧圧力が低いとき は）、高段側圧縮機 2の吸入圧力も低くなり、バイパス回路 11からバイパスされる冷媒 量が減少するので、膨張機 5によるエネルギー回収量が多くなり給湯 (又は暖房) CO P比が大きくなる。また、従来技術と比較すると、膨張機 5による回収エネルギーの減 少よりも 2段圧縮装置 3における圧縮エネルギーの減少量のほうが大きくなる結果が 、従来技術の給湯 (又は暖房) COPに対し本実施の形態の給湯 (又は暖房) COP比 が大きくなる。  Therefore, according to this refrigeration cycle apparatus, hot water heated by the radiator 4 is used for hot water supply or heating, so that hot hot water or high-temperature blown air can be obtained. In addition, the refrigeration cycle apparatus using carbon dioxide and carbon dioxide as a refrigerant can greatly improve the efficiency as compared with the prior art. FIG. 8 shows the effect of the hot water supply (or heating) COP ratio on the change in hot water temperature at the outlet of the radiator 4 (indirectly, change in high pressure) in the refrigeration cycle apparatus of the fifth embodiment. . The hot water (or heating) COP ratio in this case is defined as the hot water temperature at the condenser outlet of a hot water supply device (or heating device) that applies an old general refrigeration cycle device that does not recover energy by the expander 5. FIG. 7 shows the ratio of the COP of Embodiment 5 to the hot water supply (or heating) COP when changed. As shown in the figure, when the hot water temperature at the radiator 4 outlet is low (that is, when the high pressure is low), the suction pressure of the high-stage compressor 2 is also low, and the amount of refrigerant bypassed from the bypass circuit 11 is reduced. Therefore, the amount of energy recovered by the expander 5 increases, and the hot water supply (or heating) COP ratio increases. Compared to the conventional technology, the amount of reduction in compression energy in the two-stage compressor 3 is greater than the reduction in recovered energy by the expander 5. The hot water supply (or heating) COP ratio increases.
[0039] (第 6の実施の形態）  [0039] (Sixth embodiment)
次に、第 6の実施の形態に係る冷凍サイクル装置について図 9に従って説明する。 冷凍サイクル装置は、第 1の実施の形態と同一の冷媒回路において、膨張機 5として 非容積型のタービン型膨張機を用い、低段側圧縮機 1として非容積型のタービン型 圧縮機を採用している。  Next, a refrigeration cycle apparatus according to a sixth embodiment will be described with reference to FIG. In the refrigeration cycle apparatus, in the same refrigerant circuit as in the first embodiment, a non-displacement turbine type expander is used as the expander 5, and a non-displacement turbine type compressor is used as the low-stage compressor 1. is doing.
[0040] このような冷凍サイクルにおいては、第 1の実施の形態の冷凍サイクルと同様な効 果を奏することができる。カロえて、膨張機 5が非容積型であるため、膨張機 5の起動 負荷の軽減が実現される。また、低段側圧縮機 1としてタービン型圧縮機を用いてい るので、容積型圧縮機を用いる場合に比し、低段側圧縮機 1の起動負荷を軽減し、 惹いては、膨張機 5と低段側圧縮機 1との組み合わせ起動負荷が軽減される。 [0040] In such a refrigeration cycle, an effect similar to that of the refrigeration cycle of the first embodiment can be achieved. Since the expander 5 is non-volumetric, the start-up load of the expander 5 can be reduced. In addition, since a turbine-type compressor is used as the low-stage compressor 1, the starting load of the low-stage compressor 1 is reduced compared to the case where a positive displacement compressor is used. As a result, the combined starting load of the expander 5 and the low-stage compressor 1 is reduced.
上記各実施の形態は、以下のように変形して適用することも可能である。  Each of the above embodiments can be modified and applied as follows.
(1) 全実施の形態において、第 5の実施の形態と同様に冷媒として二酸化炭素を 使用することができる。この場合は、冷凍サイクルは超臨界サイクルとなる。  (1) In all the embodiments, carbon dioxide can be used as a refrigerant as in the fifth embodiment. In this case, the refrigeration cycle is a supercritical cycle.
(2) 第 2— 4実施の形態は、容量可変の低段側圧縮機 1として容量可変の斜板式 圧縮機を用いている力 これに限定されるものではなぐインバータ駆動のロータリ式 圧縮機など他の容量可変方法を用いた他形式の圧縮機としてもょ ヽ。  (2) In the second to fourth embodiments, the power using a variable capacity swash plate type compressor as the variable capacity low stage side compressor 1 is not limited to this, but is an inverter driven rotary compressor, etc. It can also be used as a compressor of other types using other capacity variable methods.
(3) 第 3の実施の形態においては、低段側圧縮機 1、アキュムレータ 9、膨張機 5及 び気液分離器 6がー体ィ匕されている力 これら全てを一体ィ匕する必要はない。これら のうち、任意の何れかを一体化して、機器構成及び機器接続配管の簡素化を行って ちょい。  (3) In the third embodiment, the force with which the low-stage compressor 1, the accumulator 9, the expander 5 and the gas-liquid separator 6 are combined is required to be integrated with each other. Absent. Any one of these can be integrated to simplify the equipment configuration and equipment connection piping.
(4) 第 4の実施の形態において、低段側圧縮機 1を容量可変型圧縮機とせずに、 容量固定圧縮機とし、第 1の実施の形態と同様に膨張機 33と並列に、流量制御弁を 備えたバイパス回路を設け、高段側圧縮機の吸入圧力の上昇に応じて流量制御弁 の開度を増大させるように制御してもよ 、。  (4) In the fourth embodiment, the low stage compressor 1 is not a variable capacity compressor, but a fixed capacity compressor, and in parallel with the expander 33 as in the first embodiment, the flow rate A bypass circuit having a control valve may be provided to control the opening degree of the flow control valve to increase as the suction pressure of the high-stage compressor increases.
(5) 絞り装置 7、 31としては電動膨張弁、キヤビラリ一チューブ、感温式膨張弁など 種々のものを使用することができる。  (5) As the expansion devices 7 and 31, various devices such as an electric expansion valve, a capillary tube, and a temperature-sensitive expansion valve can be used.
(6) 第 2乃至 5の実施の形態において、膨張機 5或いは低段側圧縮機 1として実施 の形態 6のような非容積型のタービン型膨張機を用いてもよい。この場合、実施の形 態 6に記載したように膨張機 5と低段側圧縮機 1との組み合わせ起動負荷を軽減する ことができる。  (6) In the second to fifth embodiments, a non-displacement type turbine expander as in the sixth embodiment may be used as the expander 5 or the low-stage compressor 1. In this case, as described in Embodiment 6, the combined starting load of the expander 5 and the low-stage compressor 1 can be reduced.
(7) 全実施の形態において、膨張機 5或いは低段側圧縮機 1として、他の非容積型 膨張機を用いてもよい。  (7) In all the embodiments, other non-volumetric expanders may be used as the expander 5 or the low-stage compressor 1.
(8) 実施の形態 1〜5において、膨張機 5或いは低段側圧縮機 1として容積型の膨 張機、例えば、ロータリー型、スクロール型、スクリュウ型、ベーン型、斜板型、バンケ ル型、ヘリカル型を用いてもよい。また、膨張機 5或いは低段側圧縮機 1として、定流 量型である歯車型、ルーツ型、ねじ型を用いてもよい。  (8) In the first to fifth embodiments, the expander 5 or the low-stage compressor 1 is a volumetric expander, for example, a rotary type, scroll type, screw type, vane type, swash plate type, or bankel type. A helical type may be used. Further, as the expander 5 or the low-stage compressor 1, a constant flow type gear type, roots type, or screw type may be used.
(9) 第 1乃至 5の実施の形態において、低段側圧縮機 1として起動時の負荷を軽減 できる起動負荷軽減装置を備えた容積型圧縮機を用い、低段側圧縮機 1の起動負 荷を軽減するようにしてもよい。このようにすれば、中小容量に最適な上述のような容 積型圧縮機を使用しながら、駆動用電動機を用いずに膨張機 5のみで低段側圧縮 機 1を起動させることが可能になる。この場合に適用し得る起動負荷軽減装置は、例 えば、起動時の回転数を減速させる方式、吸入容積を減少させる方式など一般的な ものでよい。なお、この場合において、膨張機としてタービン型等の非容積型膨張機 を用いればさらに起動負荷を軽減することができる。 (9) In the first to fifth embodiments, as the low-stage compressor 1, the load at startup is reduced. A positive displacement compressor equipped with a starting load reducing device that can be used may be used to reduce the starting load of the low-stage compressor 1. In this way, it is possible to start up the low-stage compressor 1 with only the expander 5 without using a drive motor, while using the above-described capacity compressor that is optimal for small and medium capacities. Become. The starting load reducing device applicable in this case may be a general device such as a method of decelerating the number of rotations at the time of starting, or a method of reducing the suction volume. In this case, if a non-volumetric expander such as a turbine type is used as the expander, the starting load can be further reduced.

Claims

請求の範囲 The scope of the claims
[1] 低段側圧縮機と高段側圧縮機とからなる 2段圧縮装置と、  [1] a two-stage compressor comprising a low-stage compressor and a high-stage compressor,
高圧ガス冷媒を冷却する放熱器と、  A radiator that cools the high-pressure gas refrigerant;
前記放熱器で冷却された高圧ガス冷媒を膨張させる膨張機を含む減圧装置と、前 記減圧装置により減圧された低圧冷媒を蒸発させる蒸発器と、  A decompressor including an expander that expands the high-pressure gas refrigerant cooled by the radiator; an evaporator that evaporates the low-pressure refrigerant decompressed by the decompressor;
前記減圧装置で減圧された中間圧ガス冷媒を高段側圧縮機の吸入側に導入する ガスインジェクション回路と、  A gas injection circuit for introducing the intermediate-pressure gas refrigerant decompressed by the decompression device to the suction side of the high-stage compressor;
前記膨張機と低段側圧縮機とは同軸で連結されていることと  The expander and the low-stage compressor are connected coaxially;
を特徴とする冷凍サイクル装置。  A refrigeration cycle apparatus characterized by.
[2] 請求項 1にお!、て、中間圧冷媒を気液分離する気液分離器が設けられ、前記減圧 装置は、放熱器出口側と気液分離器との間に接続されて、放熱器で冷却された後の 高圧冷媒を中間圧まで膨張させる膨張機と、気液分離器の液部分と蒸発器入口側と の間に接続された、気液分離器で気液分離された中間圧液冷媒を低圧まで減圧す る絞り装置とを有し、前記ガスインジヱクシヨン回路は、気液分離器のガス部分と高段 側圧縮機の吸入側とを連絡するように接続されてなることを特徴とする冷凍サイクル 装置。 [2] In claim 1, there is provided a gas-liquid separator for gas-liquid separation of the intermediate pressure refrigerant, and the decompression device is connected between the radiator outlet side and the gas-liquid separator, Gas-liquid separation was performed by the gas-liquid separator connected between the expander that expands the high-pressure refrigerant that has been cooled by the radiator to an intermediate pressure and the liquid part of the gas-liquid separator and the evaporator inlet side. The gas pressure circuit is connected to communicate the gas part of the gas-liquid separator and the suction side of the high-stage compressor. A refrigeration cycle apparatus characterized by comprising:
[3] 請求項 1において、高圧冷媒を流通させる高圧側経路と中間圧冷媒を流通させる 中間圧側経路とが熱交換可能に構成されてなる熱交換器が設けられ、この熱交換器 の高圧側経路の入口側は、放熱器の出口側に接続され、前記減圧装置は、放熱器 の出口側と熱交^^の中間圧側経路の入口側との間に接続された、放熱器で冷却 された後の高圧冷媒の一部を中間圧まで減圧する絞り装置と、熱交換器の高圧側経 路の出口側と蒸発器の入口側との間に接続された、放熱器及び熱交^^で冷却さ れた後の残部の高圧冷媒を低圧まで膨張させる膨張機とを有し、前記ガスインジエタ シヨン回路は、熱交換器の中間圧側経路の出口側と高段側圧縮機の吸入側との間 に接続されてなることを特徴とする冷凍サイクル装置。  [3] In Claim 1, there is provided a heat exchanger in which a high-pressure side passage through which the high-pressure refrigerant flows and an intermediate-pressure side passage through which the intermediate-pressure refrigerant flows are configured to be able to exchange heat, and the high-pressure side of the heat exchanger The inlet side of the path is connected to the outlet side of the radiator, and the pressure reducing device is cooled by a radiator connected between the outlet side of the radiator and the inlet side of the intermediate pressure side path of the heat exchanger. And a heat exchanger and a heat exchanger connected between the expansion device for reducing a part of the high-pressure refrigerant to an intermediate pressure and the outlet side of the high-pressure side passage of the heat exchanger and the inlet side of the evaporator. And an expander that expands the remaining high-pressure refrigerant after being cooled at a low pressure to a low pressure, and the gas induction circuit includes an outlet side of the intermediate pressure side path of the heat exchanger and a suction side of the high-stage compressor. A refrigeration cycle apparatus connected to each other.
[4] 請求項 1〜3の何れか 1項において、前記減圧装置は、膨張機をバイパスするバイ パス回路を有し、このバイパス回路には流量制御弁が設けられて 、ることを特徴とす る冷凍サイクル装置。 [4] In any one of claims 1 to 3, the pressure reducing device has a bypass circuit that bypasses the expander, and the bypass circuit is provided with a flow control valve. Refrigerating cycle equipment.
[5] 請求項 1〜3の何れか 1項において、前記低段側圧縮機は、容量可変型圧縮機に 構成されて!ヽることを特徴とする冷凍サイクル装置。 [5] The refrigeration cycle apparatus according to any one of claims 1 to 3, wherein the low-stage compressor is configured as a variable capacity compressor.
[6] 請求項 5において、前記低段側圧縮機は、容量可変型斜板式圧縮機であることを 特徴とする冷凍サイクル装置。 6. The refrigeration cycle apparatus according to claim 5, wherein the low-stage compressor is a variable capacity swash plate compressor.
[7] 請求項 1において、前記低段側圧縮機と膨張機とが隔壁を介して一つのケーシン グ内に収納され、さらに、低段側圧縮機の収納された室内がアキュムレータとして兼 用されていることを特徴とする冷凍サイクル装置。 [7] In Claim 1, the low-stage compressor and the expander are housed in one casing through a partition, and the chamber in which the low-stage compressor is housed is also used as an accumulator. A refrigeration cycle apparatus characterized by comprising:
[8] 請求項 2において、前記低段側圧縮機と膨張機とが隔壁を介して一つのケーシン グ内に収納され、さらに、低段側圧縮機の収納された室内がアキュムレータとして兼 用されるとともに、前記膨張機を収納する室が気液分離器として兼用されていること を特徴とするの冷凍サイクル装置。 [8] In Claim 2, the low-stage compressor and the expander are housed in one casing through a partition, and the chamber in which the low-stage compressor is housed is also used as an accumulator. And a chamber for housing the expander is also used as a gas-liquid separator.
[9] 請求項 1にお 、て、冷媒として二酸ィ匕炭素が使用されて 、ることを特徴とする冷凍 サイクル装置。 [9] A refrigeration cycle apparatus according to claim 1, wherein carbon dioxide is used as a refrigerant.
[10] 請求項 9において、低段側圧縮機の吸入側に冷媒回路内の冷媒量を調節可能と する容積に形成されたアキュムレータを備えてなることを特徴とする冷凍サイクル装 置。  [10] The refrigeration cycle device according to claim 9, further comprising an accumulator formed on a suction side of the low-stage compressor so as to adjust a refrigerant amount in the refrigerant circuit.
[11] 請求項 1において、前記膨張機は、非容積型膨張機であることを特徴とする冷凍サ イタル装置。  [11] The refrigeration sanitary apparatus according to claim 1, wherein the expander is a non-volumetric expander.
[12] 前記低段側圧縮機は、起動時の負荷を軽減できる負荷軽減装置を備えた容積型 圧縮機であることを特徴とする請求項 11記載の冷凍サイクル装置。  12. The refrigeration cycle apparatus according to claim 11, wherein the low-stage compressor is a positive displacement compressor provided with a load reducing device that can reduce a load during startup.
PCT/JP2005/014416 2004-08-06 2005-08-05 Freezing cycle apparatus WO2006013970A1 (en)

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KR101162756B1 (en) * 2007-02-24 2012-07-05 삼성전자주식회사 A water-cooling type air conditioner and control method thereof
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US20110247358A1 (en) * 2008-12-22 2011-10-13 Panasonic Corporation Refrigeration cycle apparatus
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Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS5855655A (en) * 1981-09-30 1983-04-02 株式会社東芝 Turbine for refrigerating cycle
JPS62228843A (en) * 1986-03-31 1987-10-07 株式会社東芝 Refrigeration cycle
JP2001141315A (en) * 1999-11-10 2001-05-25 Aisin Seiki Co Ltd Refrigerating air conditioner
JP2003074999A (en) * 2001-08-31 2003-03-12 Daikin Ind Ltd Refrigerating machine
JP2004101033A (en) * 2002-09-06 2004-04-02 Hoshizaki Electric Co Ltd Freezing circuit of cooling storage vessel
JP2004150749A (en) * 2002-10-31 2004-05-27 Matsushita Electric Ind Co Ltd Refrigerating cycle device
JP2004183913A (en) * 2002-11-29 2004-07-02 Mitsubishi Electric Corp Air conditioner

Patent Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS5855655A (en) * 1981-09-30 1983-04-02 株式会社東芝 Turbine for refrigerating cycle
JPS62228843A (en) * 1986-03-31 1987-10-07 株式会社東芝 Refrigeration cycle
JP2001141315A (en) * 1999-11-10 2001-05-25 Aisin Seiki Co Ltd Refrigerating air conditioner
JP2003074999A (en) * 2001-08-31 2003-03-12 Daikin Ind Ltd Refrigerating machine
JP2004101033A (en) * 2002-09-06 2004-04-02 Hoshizaki Electric Co Ltd Freezing circuit of cooling storage vessel
JP2004150749A (en) * 2002-10-31 2004-05-27 Matsushita Electric Ind Co Ltd Refrigerating cycle device
JP2004183913A (en) * 2002-11-29 2004-07-02 Mitsubishi Electric Corp Air conditioner

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