JP2007178042A - Supercritical vapor compression type refrigerating cycle and cooling and heating air conditioning facility and heat pump hot-water supply machine using it - Google Patents

Supercritical vapor compression type refrigerating cycle and cooling and heating air conditioning facility and heat pump hot-water supply machine using it Download PDF

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JP2007178042A
JP2007178042A JP2005375780A JP2005375780A JP2007178042A JP 2007178042 A JP2007178042 A JP 2007178042A JP 2005375780 A JP2005375780 A JP 2005375780A JP 2005375780 A JP2005375780 A JP 2005375780A JP 2007178042 A JP2007178042 A JP 2007178042A
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refrigerant
pressure
high
expansion device
low
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Toshihide Koda
Hideaki Maeyama
So Nomoto
Kei Sasaki
Shin Sekiya
Tetsuei Yokoyama
圭 佐々木
英明 前山
利秀 幸田
哲英 横山
宗 野本
慎 関屋
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Mitsubishi Electric Corp
三菱電機株式会社
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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
    • F25B1/00Compression machines, plant, or systems with non-reversible cycle
    • F25B1/10Compression machines, plant, or systems with non-reversible cycle with multi-stage compression
    • 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, plant, or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002Compression machines, plant, or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008Compression machines, plant, 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
    • F25B2309/00Gas cycle refrigeration machines
    • F25B2309/06Compression machines, plant or systems characterised by the refrigerant being carbon dioxide
    • F25B2309/061Compression machines, plant 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
    • 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
    • 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
    • F25B2400/23Separators
    • 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
    • F25B2600/21Refrigerant outlet evaporator temperature

Abstract

PROBLEM TO BE SOLVED: To provide a supercritical vapor compression type refrigerating cycle provided with a two-stage compressor reducing suction heating loss and having high efficiency with a heat radiation side of carbon dioxide refrigerant operating in a supercritical zone.
SOLUTION: In this supercritical vapor compression type refrigerating cycle having the two-stage compressor having an intermediate connection circuit to compress the refrigerant in two stages, a first expansion device for reducing the pressure of the refrigerant to an intermediate pressure, a gas-liquid separator for separating the refrigerant into gas and liquid, an injection circuit for injecting the gas phase side refrigerant after separating into gas and liquid into the intermediate connection circuit, and a second expansion device, an expulsion volume ratio of expulsion volume on a high stage side to expulsion volume on a low stage side is higher than the adiabatic index root of a ratio of a suction pressure of the two-stage compressor to a refrigerant saturated liquid pressure in the first expansion device. Opening or degree of suction heating of the first and second expansion devices is controlled to keep the refrigerant injected into the intermediate connection circuit in a gas phase state.
COPYRIGHT: (C)2007,JPO&INPIT

Description

この発明は、冷媒を二段圧縮するとともに中間圧に冷媒をインジェクションする二段圧縮機を備える超臨界蒸気圧縮式冷凍サイクルおよびこれを用いる冷暖房空調設備とヒートポンプ給湯機とに関するものである。 The present invention relates to an intermediate pressure using supercritical vapor compression refrigeration cycle and which comprises a two-stage compressor for injecting the refrigerant into the HVAC heat pump water heater as well as two-stage compression refrigerant.

冷暖房空調設備に用いられていた特定フロンは、オゾン層の破壊や地球温暖化などの問題があり規制されている。 Particular have been used in the HVAC CFCs, there are problems such as destruction of the ozone layer and global warming are regulated. また、新しく開発された代替冷媒は、オゾン層を破壊しないが、地球温暖化係数が二酸化炭素冷媒の数百から数千倍である。 Moreover, alternative refrigerant newly developed, but does not destroy the ozone layer, global warming coefficient is several thousand times hundreds of carbon dioxide refrigerant. このような背景から、二酸化炭素が地球環境にやさしい冷媒として再び注目されている。 Against this background, carbon dioxide is again attracting attention as a refrigerant friendly to the global environment.
しかし、二酸化炭素の臨界温度は約31℃であり、空調用冷凍サイクルの作動流体として使った場合、通常の放熱側環境温度(冷房時室外:25℃〜35℃程度、暖房時室内:15℃〜25℃程度)で圧縮された冷媒は、二酸化炭素の臨界温度と7.38MPaの臨界圧力を超えるようになり、超臨界蒸気圧縮式冷凍サイクルを形成する。 However, the critical temperature of carbon dioxide is about 31 ° C., when used as the working fluid of the air conditioning refrigeration cycle, normal radiation side ambient temperature (cooling time outdoor: 25 ° C. to 35 ° C. approximately, heating operation room: 15 ° C. refrigerant compressed at 25 approximately ° C.) is now greater than the critical pressure of the critical temperature and 7.38MPa carbon dioxide to form a supercritical vapor compression refrigeration cycle. 超臨界蒸気圧縮式冷凍サイクルの放熱過程において、二酸化炭素冷媒が超臨界圧力状態になっているため、従来冷媒のような潜熱ではなく顕熱の形で放熱が行われ、従来の冷凍サイクルに比べサイクル効率が低下してしまう問題があった。 In the heat dissipation process of a supercritical vapor compression refrigeration cycle, since the carbon dioxide refrigerant is in a supercritical pressure state, heat radiation in the form of sensible heat is performed instead of the latent heat, such as a conventional refrigerant, compared with the conventional refrigeration cycle there is a problem that the cycle efficiency is lowered.

そこで、二酸化炭素を冷媒に用いた超臨界蒸気圧縮式冷凍サイクルの効率を向上させるために、内部熱交換器とガスインジェクションとを用いたサイクルが提案されている(例えば、特許文献1参照)。 Therefore, carbon dioxide in order to improve the efficiency of supercritical vapor compression refrigeration cycle using a refrigerant cycle which uses a gas injection internal heat exchanger has been proposed (e.g., see Patent Document 1).
また、二段圧縮機の中間圧にガスインジェクションする冷凍サイクルに適した排除容積比について、HFC冷媒(R410A)、HCFC冷媒(R22)に関して検討している(例えば、特許文献2参照)。 Further, the displacement volume ratio appropriate to the refrigeration cycle for gas injection into an intermediate pressure of the two-stage compressor, is discussed with respect to HFC refrigerant (R410A), HCFC refrigerants (R22) (for example, see Patent Document 2).
また、内部中間圧型二酸化炭素冷媒ニ段ロータリ圧縮機において一段目に対する二段目の排除容積比を1対0.56〜0.8の範囲(特に0.65を推奨)に設定し、起動時の圧力変動を小さくしてオイルフォーミングが抑制できることが開示されている(例えば、特許文献3参照)。 Further, set in the internal intermediate pressure type carbon dioxide refrigerant two-stage rotary compressor displacement volume ratio of the second stage for the first stage in the range of 1-to-0.56 to 0.8 (especially recommended 0.65), when starting smaller to the oil forming the pressure fluctuations is disclosed can be suppressed (for example, see Patent Document 3).

特開2001−116376号公報 JP 2001-116376 JP 特開2000−87892号公報 JP 2000-87892 JP 特開2001−73976号公報 JP 2001-73976 JP

しかし、従来提案では中間圧にガスインジェクションする冷凍サイクルであるが、高圧側が臨界動作する二酸化炭素冷媒を用いた場合については考慮されていない。 However, although the conventional proposed a refrigeration cycle for gas injection into an intermediate pressure, does not consider the case of using the carbon dioxide refrigerant high-pressure side is operated critical. また、別の従来提案では内部中間圧型二段ロータリ圧縮機で二酸化炭素冷媒を用いた冷凍サイクルであるが、中間圧にガスインジェクションする場合に適した排除容積比については検討していない。 Although in another previously proposed a refrigeration cycle using carbon dioxide refrigerant inside the intermediate pressure type two-stage rotary compressor, not studied displacement volume ratio appropriate for gas injection into an intermediate pressure.
二酸化炭素を冷媒に用いた超臨界蒸気圧縮式冷凍サイクルにおいては、高圧放熱器で超臨界状態に達するため、従来の提案に従って二段圧縮機の排除容積比を決定しても、気液分離が適切に行なえず中間インジェクションによる性能改善効果と信頼性向上効果とが十分に得られないという問題がある。 In the supercritical vapor compression refrigeration cycle using carbon dioxide refrigerant, to reach the supercritical state at high pressure radiator, be determined displacement volume ratio of the two-stage compressor according to the conventional proposal, the gas-liquid separation there is a problem that the performance improvement effect by the intermediate injection and reliability improvement can not be sufficiently obtained without performing properly.

この発明の目的は、吸入加熱損失が小さく高効率の高圧シェル型の二段圧縮機を備え、二酸化炭素冷媒の放熱側が超臨界域で動作する超臨界蒸気圧縮式冷凍サイクルを提供することである。 The purpose of this invention, suction heat loss is small with a two-stage compressor of the high pressure shell type high efficiency, is that the heat radiation side of the carbon dioxide refrigerant to provide a supercritical vapor compression refrigeration cycle that operates in a supercritical range .

この発明に係わる超臨界蒸気圧縮式冷凍サイクルは、低圧の冷媒が低段側回転圧縮要素で中間圧まで圧縮された冷媒が中間連結回路を経由して高段側回転圧縮要素に吸入され、上記高段側回転圧縮要素で高圧まで圧縮される二段圧縮機と、高圧に圧縮される冷媒を飽和液圧以下まで減圧する第1の膨張装置と、飽和液圧以下で湿りガス状態になった冷媒を気液分離する気液分離器と、気液分離後の気相側冷媒を上記中間連結回路にインジェクションするインジェクション回路と、気液分離後の液相側冷媒を低圧まで減圧する第2の膨張装置と、低圧まで減圧される冷媒を蒸発させる蒸発器と、を有する超臨界蒸気圧縮式冷凍サイクルにおいて、上記高段側回転圧縮要素の排除容積の上記低段側回転圧縮要素の排除容積に対する排除容積比は、 The supercritical vapor compression refrigeration cycle according to the invention is sucked refrigerant pressure refrigerant compressed by the low stage side rotary compressing element to an intermediate pressure via an intermediate connecting circuit high stage side rotary compressing element, the a two-stage compressor in the high stage side rotary compressing element is compressed to a high pressure, first an expansion device for reducing the pressure of the refrigerant compressed to a high pressure to under saturated liquid pressure or became wet gas state under the saturated liquid pressure or a gas-liquid separator for gas-liquid separation of the refrigerant, the gas-liquid and injection circuit for injecting a gas-phase refrigerant after separation in the intermediate connecting circuit, a second decompressing the liquid-phase refrigerant after gas-liquid separation to a low pressure an expansion device, an evaporator for evaporating the refrigerant to be reduced to a low pressure, the supercritical vapor compression refrigeration cycle having, for displacement volume of the low stage side rotary compressing element of the displacement volume of the high stage side rotary compressing element exclusion volume ratio, 記二段圧縮機の吸入圧力の上記第1の膨張装置における冷媒飽和液圧に対する比の等エントロピ指数の乗根以上であり、上記第1の膨張装置または上記第2の膨張装置の開度または吸入加熱度を、上記中間連結回路にインジェクションする冷媒を気相状態に保つように制御する。 Serial not less than root of isentropic exponent ratio refrigerant saturated liquid pressure in the first expansion device of the suction pressure of the two-stage compressor, the opening degree of the first expansion device or the second expansion device or the suction superheat is controlled so as to maintain the refrigerant is injected to the intermediate connecting circuit to the gas phase.

この発明に係わる超臨界蒸気圧縮式冷凍サイクルの効果は、二段圧縮機に中間インジェクションする冷媒主成分を気相状態に保つことができ、液相冷媒を多量注入することによって起こりうる圧縮機効率の低下と信頼性の低下を防ぐことができる。 The effect of the supercritical vapor compression refrigeration cycle according to the present invention, can keep refrigerant main component intermediate injection to two-stage compressor in gaseous state, the compressor efficiency can occur by large amounts injecting liquid refrigerant it is possible to prevent the lowering of the decrease in reliability.

実施の形態1. The first embodiment.
図1は、この発明の実施の形態1に係わる気液分離器を用いた超臨界蒸気圧縮式冷凍サイクルの回路図である。 Figure 1 is a circuit diagram of a supercritical vapor compression refrigeration cycle using a gas-liquid separator according to the first embodiment of the present invention.
この発明の実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルは、冷媒を二段で圧縮する二段圧縮機1と冷媒を二段で膨張する第1の膨張装置3および第2の膨張装置7とを備える二段圧縮二段膨張方式を採用している。 The supercritical vapor compression refrigeration cycle according to a first embodiment of the invention, the first expansion device 3 and the second expansion device for expanding a two-stage compressor 1 and the refrigerant for compressing a refrigerant in two stages with two-stage It employs a two-stage compression two-stage expansion system and a 7. 冷媒としては、二酸化炭素を用いている。 The refrigerant, carbon dioxide is used.
実施の形態1に係わる二段圧縮機1では、冷媒を密閉容器13に内包される低段側回転圧縮要素11および高段側回転圧縮要素12で順に圧縮して超臨界状態まで昇圧し、高段側回転圧縮要素12の吐出口d から密閉容器13内に吐出する。 In two-stage compressor 1 according to the first embodiment, in the low stage side rotary compressing element 11 and the high stage side rotary compressing element 12 is enclosing the refrigerant into the sealed container 13 is compressed in order to boost up the supercritical state, high It discharges into the sealed container 13 from the discharge port d 2 of the stage side rotary compressing element 12. その後、密閉容器13から外部回路へ吐出された冷媒は、高圧放熱器2で放熱冷却された後、第1の膨張装置3で飽和圧力以下まで減圧し、湿りガス状態になった冷媒を気液分離器4で気液分離する。 Thereafter, the refrigerant discharged from the sealed container 13 to an external circuit, after being radiated cooled by high-pressure radiator 2, the pressure was reduced to saturation pressure below the first expansion device 3, a gas-liquid refrigerant becomes wet gas state the gas-liquid separation in the separator 4. 低段側回転圧縮要素11の吐出口d と高段側回転圧縮要素12の吸入口S が中間連結回路15により連結されている。 Inlet S 2 of the discharge port d 1 and high stage side rotary compressing element 12 of the low stage side rotary compressing element 11 are connected by an intermediate connecting circuit 15.

気液分離器4内の気相を主成分とする冷媒は、インジェクション回路5から二段圧縮機1の中間連結回路15の途中にある冷媒混合器14に中間インジェクションされ、低段側回転圧縮要素11から吐出された冷媒に混合されてから高段側回転圧縮要素12に吸入される。 Refrigerant mainly composed of vapor of the gas-liquid separator 4 is intermediate injection the refrigerant mixer 14 on the way from the injection circuit 5 of the intermediate connecting circuit 15 of the two-stage compressor 1, a low stage side rotary compressing element 11 is sucked into the high stage side rotary compressing element 12 from being mixed in the refrigerant discharged from.
インジェクション回路5には、流量調整弁16が介設されており、インジェクション量を調整する。 The injection circuit 5, a flow rate regulating valve 16 is interposed, to adjust the injection amount.
一方、気液分離器4内の液相冷媒は、第2の膨張装置7でさらに減圧され、蒸発器8で吸熱加熱され気相状態まで蒸発されて、再び二段圧縮機1の低段側回転圧縮要素11の吸入口S から吸入される。 On the other hand, the liquid phase refrigerant in the gas-liquid separator 4 is further reduced in pressure by the second expansion device 7, is evaporated to gas phase is absorbed heated by the evaporator 8, again the lower stage of the two-stage compressor 1 It is sucked from the suction port S 1 rotary compression element 11.

また、二段圧縮機1において、中間圧が冷媒の臨界圧力以下になるように、低段側回転圧縮要素11の排除容積に対する高段側回転圧縮要素12の排除容積の割合(以下、排除容積比と称す。)は、二段圧縮機1の吸入圧力を第1膨脹装置3における冷媒飽和液圧で除算した商の等エントロピ指数乗根以上である。 Further, in the two-stage compressor 1, so that the intermediate pressure is below the critical pressure of the refrigerant, the ratio of displacement volume of the high stage side rotary compressing element 12 for displacement volume of the low stage side rotary compressing element 11 (hereinafter, displacement volume referred to as a ratio.) is the suction pressure of the two-stage compressor 1 first expander isentropic exponent root or more items divided by the refrigerant saturation pressure at 3.

また、実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルは、温度計20、21が第1の膨張装置3の上流側と下流側に、温度計22が気液分離器4に連なる液相側配管23に、温度計24が気液分離器4に連なるインジェクション回路5に、温度計25、26が低段側回転圧縮要素11の吐出口d から高段側回転圧縮要素12の吸入口S とを連結する中間連結回路15のインジェクション回路5が接続される位置の前後に、温度計27、28が高圧放熱器2の前後に、温度計29、30が蒸発器8の前後にそれぞれ取り付けられ、測定した温度から冷媒の状態を推定する。 Furthermore, the supercritical vapor compression refrigeration cycle according to the first embodiment, the thermometer 20 and 21 upstream and downstream of the first expansion device 3, liquid phase thermometer 22 is connected to the gas-liquid separator 4 the side pipe 23, the injection circuit 5 which thermometer 24 is connected to the gas-liquid separator 4, inlet temperature gauge 25 and 26 the high stage side rotary compressing element 12 from the discharge port d 1 of the low-stage side rotary compressing element 11 before and after the position where the injection circuit 5 of the intermediate connecting circuit 15 for coupling the S 2 is connected, in the longitudinal thermometer 27 and 28 of the high-pressure radiator 2, thermometer 29 and 30 respectively before and after the evaporator 8 attached to estimate the state of the refrigerant from the measured temperature. そして、気液分離器4の吸入前の冷媒が飽和液圧以下で、湿りガス状態にするために、気液分離器4の吸入前の温度計21の温度から飽和液温度以下、すなわち湿りガス状態であるか否かを判断し、気液分離器4に吸入される直前の冷媒の温度が液温度帯であれば、第1の膨脹装置3を絞って中間圧を下げるように調節する。 The gas-liquid separator 4 the refrigerant before inhalation under saturated liquid pressure of, in order to wet gas state, the saturated liquid temperature from the gas-liquid thermometer 21 temperature prior to inhalation of the separator 4 or less, i.e. humid gas it is determined whether the state, the temperature of the refrigerant just before it is sucked into the gas-liquid separator 4 is as long as the liquid temperature zone is adjusted to decrease the intermediate pressure squeezing first expander 3.

図2は、この発明の実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルのP−h線図である。 Figure 2 is a P-h diagram of a supercritical vapor compression refrigeration cycle according to the first embodiment of the present invention.
次に、中間インジェクションを行う効果について図2を参照して説明する。 Next, the effect of performing the intermediate injection with reference to FIG. 2 will be described. なお、二段圧縮機1の低段側吸入圧力をP S1 (二段圧縮機1の吸入圧力P SSに等しい)、中間圧力をP (低段側吐出圧力P d1と高段側吸入圧力P S2とに等しい)、高段側吐出圧力をP d2 (二段圧縮機1の吐出圧力P に等しい)で表す。 Incidentally, (equal to the suction pressure P SS of the two-stage compressor 1) a low-stage suction pressure of the two-stage compressor 1 P S1, the intermediate pressure P m (low-stage discharge pressure P d1 and high-stage suction pressure equal to the P S2), representative of the high-stage discharge pressure P d2 (equal to the discharge pressure P d of the two-stage compressor 1). また、蒸発器8を循環する循環量をG eV 、中間インジェクション量をG inj 、ガスクーラ側循環流量をG gCで表す。 Further, representative of the amount of circulating G eV circulating evaporator 8, intermediate injection amount G inj, a gas cooler circulation flow rate G gC. ここで、ガスクーラ側循環流量G gCは、G gC =G eV +G injから求まり、中間インジェクション量の割合αを、α=G inj /G eVで定義する。 Here, gas cooler circulation flow rate G gC are obtained in decreasing G gC = G eV + G inj , the ratio of the intermediate injection quantity alpha, defined by α = G inj / G eV.

二段圧縮における中間インジェクションによる成績係数(Coefficient of Performanceの略語COPと称す。)の改善は、圧縮機効率とサイクル効率とが改善される。 Improvement of coefficient of performance by intermediate injection (referred to as the abbreviation COP of Coefficient of Performance.) In a two-stage compression, the compressor efficiency and cycle efficiency can be improved.
まず、サイクル効率の改善効果について説明する。 First described the effect of improving the cycle efficiency.
(1)空調冷房用途では、二段圧縮インジェクションによって凝縮器側のエンタルピ差が増加し、冷房能力と冷房成績係数が向上する。 (1) In the air conditioning cooling applications, increased enthalpy difference of condenser side by two-stage compression injection, cooling capacity and cooling coefficient of performance is improved.
(2)空調暖房用途では、ガスクーラ側冷媒循環量が増加して暖房能力と成績係数が向上するが、空調冷房用途に比べると効果は小さい。 (2) In the air conditioning heating applications, although gas cooler refrigerant circulation amount is improved heating capacity and coefficient of performance increases, the effect is small compared to the air-conditioning cooling applications.
(3)給湯用途においても空調暖房用途と同様の効果があるが、吐出温度が下がるためインジェクション量が制約される。 (3) Although the effect is as air-conditioning heating applications even in hot water applications, the injection amount since the discharge temperature drops is restricted.
(4)高圧シェル型圧縮機を高圧縮比で運転すると、高段側吐出温度が異常に高温となり二段圧縮機1の信頼性が損なわれる。 (4) When operating at high pressure shell type compressor high compression ratio, high-stage side discharge temperature abnormally reliability of the two-stage compressor 1 becomes a high temperature is impaired. この対策として中間インジェクションすることによって、成績係数を低下させずに高段側吐出温度の異常上昇を抑えることができる。 By intermediate injection as a countermeasure, it is possible to suppress the abnormal rise of the high-stage side discharge temperature without reducing the coefficient of performance.

図3は、この発明の実施の形態1に係わる二段圧縮機1の高段側回転圧縮要素12の吸入口S でのP−h線図を示す。 Figure 3 illustrates a P-h diagram of the suction port S 2 of the high-stage rotary compression element 12 of the embodiment two-stage compressor 1 according to the first embodiment of the present invention. なお、図3において、A は、インジェクション無しの理想状態、すなわち、体積効率が1で、吸入加熱損失がない場合、A は、インジェクション有りの理想状態、すなわち、体積効率が1で、吸入加熱損失がない場合、B は、インジェクション無しで体積効率が実体積効率、吸入加熱損失がない場合、B は、インジェクション有りで体積効率が実体積効率、吸入加熱損失がない場合、C は、インジェクション無しの実状態、すなわち、体積効率が実体積効率、吸入加熱損失がある場合、C は、インジェクション有りの実状態、すなわち、体積効率が実体積効率、吸入加熱損失がある場合のそれぞれの高段側回転圧縮要素12の吸入口S におけるエンタルピと圧力を示す。 Incidentally, in FIG. 3, A 0 is an ideal state without injection, i.e., the volume efficiency is 1, if there is no suction heat loss, A 1 is injection there of the ideal state, i.e., the volume efficiency is 1, the suction If no heating loss, B 0 is the volumetric efficiency is the actual volumetric efficiency without injection, when there is no suction heat loss, B 1 is volumetric efficiency actual volume efficiency there injection, when there is no suction heat loss, C 0 the real state without injection, i.e., the volumetric efficiency is the actual volumetric efficiency, if there is a suction heat loss, C 1 is the actual state of there injection, i.e., the volumetric efficiency is the actual volumetric efficiency, in the case where there is a suction heat loss shows the enthalpy and pressure at the suction port S 2 of each of the high-stage side rotary compressing element 12.
また、中間圧は、インジェクション有りの方が無しの場合より大きい。 Greater than In addition, intermediate-pressure, when there is no better of there injection. また、実状態の中間圧の方が理想状態の中間圧より大きい。 Further, towards the intermediate pressure of the real state is higher than the intermediate pressure in the ideal state.

次に、二段圧縮インジェクションサイクルが成立するために必要な排除容積比の条件について説明する。 Next, a description will be given condition displacement volume ratio necessary for two-stage compression injection cycle is established.
1)インジェクションが行われないときの理想状態、すなわち、体積効率が1で中間加熱が行われていないときについて説明する。 1) the ideal state when the injection is not performed, i.e., will be described when the volumetric efficiency intermediate heating is not performed at 1. (図3のA の状態のときである。) (It is the state of A 0 in FIG. 3.)
このとき、理想的な中間圧P adは、低段側回転圧縮要素11において等エントロピ圧縮されたときの低段側吐出圧力P d1 adに等しく、式(1)により求められる。 In this case, an ideal intermediate pressure P m ad is equal to the low-stage discharge pressure P d1 ad when it is isentropically compressed in the low stage side rotary compressing element 11 is calculated by Equation (1). 但し、V St1は低段側排除容積、V St2は高段側排除容積、ρ S1は低段側の吸入冷媒密度、ρ d1adは等エントロピ圧縮されたときの低段側吐出冷媒密度、P S1は二段圧縮機1の低段側吸入圧力、P は二段圧縮機1の中間圧力、nは等エントロピ指数である。 However, V St1 is low-stage displacement volume, V St2 is the high-stage displacement volume, the suction refrigerant density of the [rho S1 is low-stage side, [rho D1ad the low-stage discharge refrigerant density when being isentropically compressed, P S1 the low-stage suction pressure of the two-stage compressor 1, P m is the intermediate pressure of the two-stage compressor 1, n is an isentropic exponent.

ad =P d1 ad =P S1 ×(ρ d1ad /ρ S1 P m ad = P d1 ad = P S1 × (ρ d1ad / ρ S1) n
=P S1 ×(V St1 /V St2・・・(1) = P S1 × (V St1 / V St2) n ··· (1)

等エントロピ指数nは、REFPROP第7版のデータベースから、n=LN(P /P S1 )÷LN(ρ /ρ S1 )を用いて計算することにより求められ、等エントロピ指数nは約1.3となる。 Isentropic exponent n from REFPROP 7th Edition database, obtained by calculating with n = LN (P m / P S1) ÷ LN (ρ m / ρ S1), isentropic exponent n is about 1 a .3.
また、二段圧縮機1の高段側吸入圧力P S2は、理想的な中間圧P adに等しい。 Further, the high-stage suction pressure P S2 of the two-stage compressor 1 is equal to the ideal intermediate pressure P m ad.

2)次に、インジェクションが行われ、体積効率が実体積効率で、且つ中間加熱が行われないときについて説明する。 2) Then, the injection is performed, the volumetric efficiency is a real volumetric efficiency, and will be described when the intermediate heating is not performed. (図3のB の状態のときである。) (It is the state of B 1 in FIG.)
このとき、質量保存式は式(2)で表される。 In this case, the mass conservation equation is expressed by Equation (2). 但し、η V1は低段側体積効率、η V2は高段側体積効率、ρ S2は高段側の吸入冷媒密度である。 However, eta V1 is low-stage volumetric efficiency, eta V2 is higher stage volumetric efficiency, the [rho S2 is suction refrigerant density of the high-stage side.

St1 ×η V1 ×ρ S1 ×(1+α)=V St2 ×η V2 ×ρ S2・・・(2) V St1 × η V1 × ρ S1 × (1 + α) = V St2 × η V2 × ρ S2 ··· (2)

インジェクションされた冷媒が低段側吐出口d から高段側吸入口S まで等圧変化するとき、混合した冷媒のエネルギーが保存され、式(3)が成り立つ。 When injected refrigerant is equal pressure change from the low-stage discharge port d 1 to the high-stage suction port S 2, the energy of the mixed refrigerant is stored, Equation (3) it holds. そして、排除容積比V St2 /V St1は、式(3)を式(2)に代入して得られる式(4)で求められる。 The displacement volume ratio V St2 / V St1 is determined by the formula obtained by substituting equation (3) into equation (2) (4).

ρ S2 ={(1+α)×ρ d1 ×ρ inj }/(ρ d1 ×α+ρ inj ) ・・・(3) ρ S2 = {(1 + α ) × ρ d1 × ρ inj} / (ρ d1 × α + ρ inj) ··· (3)
St2 /V St1 =(η V1 /η V2 )×(1+α)×(ρ S1 /ρ S2 V St2 / V St1 = (η V1 / η V2) × (1 + α) × (ρ S1 / ρ S2)
=(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(ρ S1 /ρ d1 ) ・・・(4) = (Η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (ρ S1 / ρ d1) ··· (4)

そして、気液分離後にガスインジェクションすることができるための必要条件は、理想状態の中間圧力P adが第1の膨張装置3の飽和液圧力P liqより小さいことである。 The prerequisite for that can be gas injection after the gas-liquid separation is that the intermediate pressure P m ad of the ideal state is less than the saturated liquid pressure P liq the first expansion device 3. この必要条件を書き直すと、式(5)、さらに式(6)となる。 Rewriting this requirement, the formula (5), further (6). ここで、低段側吐出口d が理想的低段側吐出口d adに等しいとすると、必要条件は式(7)となる。 Here, the when the low-stage discharge port d 1 is equal to the ideal low-stage discharge port d 1 ad, requirement formula (7).

S1 ×(ρ d1 ad /ρ S1 <P liq <P crt・・・(5) P S1 × (ρ d1 ad / ρ S1) n <P liq <P crt ··· (5)
ρ S1 /ρd ad >(P s1 /P liq1/n >(P s1 /P crt1/n・・・(6) ρ S1 / ρd 1 ad> ( P s1 / P liq) 1 / n> (P s1 / P crt) 1 / n ··· (6)
ρ S1 /ρd >(P s1 /P liq1/n >(P s1 /P crt1/n・・・(7) ρ S1 / ρd 1> (P s1 / P liq) 1 / n> (P s1 / P crt) 1 / n ··· (7)

ゆえに、排除容積比V St2 /V St1についての必要条件は、式(4)と式(7)から式(8)となる。 Thus, requirements for displacement volume ratio V St2 / V St1 becomes Equation (4) from equation (7) and (8).

St2 /V St1 >(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(P s1 /P liq1/n V St2 / V St1> (η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (P s1 / P liq) 1 / n
>(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(P s1 /P crt1/n・・・(8) > (Η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (P s1 / P crt) 1 / n ··· (8)

通常、ロータリ式の回転圧縮要素を同じシリンダ(内径と厚みが等しい)を用いて構成すると、漏れ感度(=漏れ面積÷排除容積)は排除容積が小さいほど大きくなることが知られている。 Usually, when configured using a rotary compression element of the rotary type same cylinder (inner diameter and the thickness are equal), the leakage sensitivity (= leakage area ÷ displacement volume) is known to increase as the displacement volume is small. 通常、低段側排除容積V St1が高段側排除容積V St2より大きいので、(すなわち、ここでは、V St2 /V St1 <1とすると、)低段側体積効率η V1が高段側体積効率η V2より大きくて、式(9)の関係が成り立つ。 Usually, since the low-stage displacement volume V St1 is greater than the high-stage displacement volume V St2, (i.e., here, if the V St2 / V St1 <1, ) the low-stage volumetric efficiency eta V1 is higher stage volume greater than efficiency eta V2, the relationship of formula (9) is satisfied. また、式(4)の(1+α×ρ d1 /ρ inj )に関して式(10)の関係が成り立つので、排除容積比V St2 /V St1についての必要条件は、式(11)となる。 Further, the relation of formula (10) with respect to the formula (4) (1 + α × ρ d1 / ρ inj) is satisfied, requirements for displacement volume ratio V St2 / V St1 becomes equation (11).

η V1 /η V2 >1 ・・・(9) η V1 / η V2> 1 ··· (9)
(1+α×ρ d1 /ρ inj )>1 ・・・(10) (1 + α × ρ d1 / ρ inj)> 1 ··· (10)
St2 /V St1 >(P s1 /P liq1/n >(P s1 /P crt1/n・・・(11) V St2 / V St1> (P s1 / P liq) 1 / n> (P s1 / P crt) 1 / n ··· (11)

3)次に、インジェクションが行われ、体積効率が実体積効率で、且つ中間加熱が行われているときについて説明する。 3) Then, the injection is performed, the volumetric efficiency is a real volumetric efficiency, and will be described when the intermediate heating is being performed. (図3のC の状態のときである。) (It is the state of C 1 in FIG.)
各種圧縮機損失の発生により、中間圧の冷媒は等圧のまま過熱され温度上昇すると、吐出される中間圧の冷媒密度ρ d1 は、低段側回転圧縮要素11で等エントロピ圧縮された後での低段側吐出冷媒密度ρ d1 adより大きくなり、式(12)の関係式が成り立つ。 The occurrence of various compressor losses, the intermediate pressure refrigerant is superheated remain isobaric temperature rises, the refrigerant density [rho d1 r of intermediate pressure discharged after being isentropically compressed in the low stage side rotary compressing element 11 greater than the low-stage discharge refrigerant density [rho d1 ad in relation expression of expression (12) holds. また、式(4)のρ d1にρ d1 を代入すると、式(13)が得られる。 Also, substituting [rho d1 r in [rho d1 of formula (4), equation (13) is obtained. この式(13)のρ d1 をρ d1 adで置き換え、関係式(12)から式(14)が得られる。 The [rho d1 r of the equation (13) replaced with [rho d1 ad, equation (14) is obtained from equation (12).

ρ d1 ad <ρ d1 ・・・(12) ρ d1 ad <ρ d1 r ··· (12)
St2 /V St1 =(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(ρ S1 /ρ d1 ) ・・・(13) V St2 / V St1 = (η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (ρ S1 / ρ d1 r) ··· (13)
St2 /V St1 >(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(ρ S1 /ρ d1 ad V St2 / V St1> (η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (ρ S1 / ρ d1 ad)
=(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(P S1 /P ad ) ・・・(14) = (Η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (P S1 / P m ad) ··· (14)

なお、気液分離後ガスインジェクションを行えるための排除容積比に関する必要条件は、式(8)、さらに式(11)となる。 Note that requirements for displacement volume ratio for enabling the gas injection after the gas-liquid separation, the formula (8), and further equation (11).

4)図3のA の状態のときは、式(8)のαを零とし、η V1 =η V2とすればよい。 4) the state of A 0 in FIG. 3, the α of formula (8) is set to zero, or if η V1 = η V2. 図3のB の状態のときは、式(8)のαを零とし、ρ d1 ad =ρ d1 とすればよい。 The state of the B 0 in FIG. 3, the zero α of the equation (8) may be set to ρ d1 ad = ρ d1 r. 図3のC の状態のときは、式(8)のαを零とし、ρ d1 ad <ρ d1 とすればよい。 In the state of C 0 in FIG. 3, the zero α of the equation (8) may be set to ρ d1 add1 r.
また、図3のA の状態のときは、式(8)のα>0とし、η V1 =η V2とすればよい。 Further, when the state of the A 1 in FIG. 3, and alpha> 0 of the formula (8) may be set to η V1 = η V2.
また、気液分離後ガスインジェクションを行えるための排除容積比に関する必要条件は、式(8)、さらに式(11)となる。 Moreover, requirements for displacement volume ratio for enabling the gas injection after the gas-liquid separation, the formula (8), and further equation (11).

次に、実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルの効果について説明する。 Next, a description will be given of an effect of the supercritical vapor compression refrigeration cycle according to the first embodiment.
表1には、住宅用二酸化炭素冷媒ヒートポンプ給湯機に用いられる代表的な環境条件と運転条件を示す。 Table 1 shows the typical environmental conditions and operating conditions used in residential carbon dioxide refrigerant heat pump water heater. この環境条件は、日本冷凍空調工業会標準規格「JRA4050−2005」に記載される温度条件および(財)ベターリビング制定の優良住宅部品性能試験方法書(BLT EH:2003)で定められた値から引用した。 The environmental conditions, Japan Refrigeration and Air Conditioning Industry Association standard "JRA4050-2005" temperature conditions and is described in (goods) Better Living enactment of prime residential parts performance test method statement: quote from the value determined by (BLT EH 2003) did. SHはSuper−Heatの略で吸入加熱度を表し、ここでは約10℃を仮定した。 SH represents the suction superheat stands for Super-Heat, where it was assumed about 10 ° C.. 添字Sは吸入、dは吐出、expは膨張弁前を意味する。 Subscript S inhalation, d is discharge, exp denotes the pre-expansion valve. 運転条件に示す吸入、吐出、膨張弁前の温度と圧力は、環境条件温度から現実的な範囲でほぼ一様に定められる。 Inhalation shown in operating conditions, discharge, temperature and pressure before expansion valve is determined almost uniformly in a realistic range of environmental conditions temperature.

表2には、実施の形態1の効果を比較するために用いるロータリ式の二段圧縮機1の基本仕様を示す。 Table 2 shows the basic specifications of the two-stage compressor 1 for rotary used to compare the effects of the first embodiment.

また、表3には、二段圧縮機1の性能特性のうち、中間圧の予測値と実験値を示す。 Further, Table 3, among the performance characteristics of the two-stage compressor 1, showing the predicted value of the intermediate pressure between the experimental values. なお、予測方法は、技術文献1(福田充宏、他3名、「R410A用2段ロータリ圧縮機の性能予測」、平成10年度日本冷凍空調学会学術講演会講演論文集、日本冷凍空調学会、平成10年10月、p.41−44)や技術文献2(角田昌之、他1名、「回転式容積形圧縮機の内部漏れの解析と評価」、第19回空気調和・冷凍連合講演会講演論文集、日本冷凍空調学会、1985年4月、p.17−19)に記載されている解析手段を適用する。 It should be noted that the prediction method, the technical literature 1 (Mitsuhiro Fukuda, and three others, "performance prediction of the two-stage rotary compressor for R410A", 1998 Japan Society of Refrigerating and Air Conditioning Engineers academic lecture Proceedings, Japan Society of Refrigerating and Air Conditioning Engineers, Heisei 10 October, p.41-44) and technical literature 2 (Masayuki Tsunoda, and one other person, "evaluation and analysis of internal leakage of rotary positive displacement compressor", the 19th air-conditioning and refrigeration Association lecture lecture Collected papers, Japan Society of refrigerating and Air Conditioning Engineers, April 1985, to apply the analysis means that are described in p.17-19).

環境条件が給湯定格およびJRAIA冬(冬季高温加熱)において、インジェクション量の割合(α)が10%のときにインジェクションを行った場合と行わない場合とで中間圧を予測すると、表3に示すように、理想状態(A )の中間圧予測値より実体積効率を考慮した状態(B )の中間圧予測値のほうが大きく、さらに、吸入加熱損失を考慮した状態(C )の中間圧予測値のほうが大きくなり、実測値とよく一致することが確認できた。 In environmental conditions hot water rating and JRAIA Winter (Winter high-temperature heating), when the ratio of the injection amount (alpha) predicts an intermediate pressure in the case of not performing the case of performing injection at 10%, as shown in Table 3 the intermediate pressure state intermediate pressure considering than expected actual volumetric efficiency (B 1) is increased towards the intermediate pressure predictive value, further conditions in consideration of the intake heating loss (C 1) of the ideal state (a 1) more of the predicted value is large, it was confirmed that good agreement with measured values.

次に、気液分離式の二段圧縮ガスインジェクションサイクルにおいて、二段圧縮機1の排除容積比が0.65と0.85との場合の中間インジェクションによる性能改善効果を予測し比較した結果を説明する。 Then, the two-stage compression gas injection cycle of the gas-liquid separating type, the results of performance improvement compared predicted by the intermediate injection when the displacement volume ratio of the two-stage compressor 1 is 0.65 and 0.85 explain. 予測するときの計算条件として、表1に示す給湯定格の運転条件を用いている。 As calculation conditions for prediction, it is used operating conditions of the hot water supply rating shown in Table 1. また、インジェクション量の割合は、気液分離後の気相ガスが最大限にガスインジェクションする場合を仮定した。 The ratio of injection amount, vapor gas after gas-liquid separation is assumed a case where the gas injection maximally.

図4は、排除容積比およびインジェクションの有無をパラメータとする全圧縮比P d2 /P S1に対する圧縮機効率の関係を示すグラフである。 Figure 4 is a graph showing the relationship between the compressor efficiency for the entire compression ratio P d2 / P S1 of the presence or absence of displacement volume ratio and injection as parameters. 図5は、排除容積比およびインジェクションの有無をパラメータとする全圧縮比P d2 /P S1に対する給湯加熱成績係数(COP)の関係を示すグラフである。 Figure 5 is a graph showing the relationship between the hot water heating coefficient of performance (COP) to the total compression ratio P d2 / P S1 of the presence or absence of displacement volume ratio and injection as parameters.
図4と図5から分かるように、インジェクション無しの場合には、排除容積比0.85の二段圧縮機1は全圧縮比2.5以下で圧縮機効率および給湯加熱成績係数などの性能が優れており、排除容積比0.65の二段圧縮機1は全圧縮比2.5より大きい領域で性能が優れている。 As can be seen from FIGS. 4 and 5, in the case without injection is two-stage compressor 1 of displacement volume ratio 0.85 performance such as the compressor efficiency and hot water heating coefficient of performance with less total compression ratio of 2.5 excellent, two-stage compressor 1 of displacement volume ratio 0.65 has excellent performance in the entire compression ratio greater than 2.5 regions. 一方、インジェクション有りの場合には、全圧縮比が約2から約4の範囲で排除容積比0.85の二段圧縮機1の性能が優れている。 On the other hand, in the case of there injection, the total compression ratio is better bunk performance of the compressor 1 of the displacement volume ratio 0.85 in the range of about 2 to about 4.

図6は、排除容積比をパラメータとする全圧縮比に対するガスインジェクション量の割合の関係を示すグラフである。 Figure 6 is a graph showing the relationship between the proportion of the gas injection amount to the total compression ratio of a displacement volume ratio as a parameter.
気液分離式ガスインジェクションをするためには、第1膨脹装置3で減圧された冷媒が飽和液圧P liq以下で、湿りガス状態になることが必要条件であり、乾き度βが大きいほどインジェクション量の割合αが大きくなる。 To the gas-liquid separation type gas injection is a refrigerant decompressed by the first expansion device 3 is less than the saturation pressure P liq, is it necessary condition that the wet gas state, the larger the dryness β Injection ratio of the amount α increases. ところが、排除容積比0.65の場合には、中間圧が比較的高いので、全圧縮比2.5以下で気液分離できない。 However, in the case of displacement volume ratio 0.65, the intermediate pressure is relatively high, it can not be gas-liquid separation in all compression ratio of 2.5 or less. 全圧縮比2.5を超えても乾き度βが小さいので大きなインジェクション量は得られない。 No significant injection amount can be obtained because the total compression ratio of 2.5 is also the dryness β is smaller than a.

図7は、排除容積比0.85の二段圧縮機1におけるインジェクションの有無をパラメータとする全圧縮比に対する低段側圧縮比および高段側圧縮比の関係を示すグラフである。 Figure 7 is a graph showing the relationship between the low-stage side compression ratio and high-stage compression ratio to total compression ratio to the presence of the injection parameters of the two-stage compressor 1 of displacement volume ratio 0.85. 図8は、排除容積比0.65の二段圧縮機1におけるインジェクションの有無をパラメータとする全圧縮比に対する低段側圧縮比および高段側圧縮比の関係を示すグラフである。 Figure 8 is a graph showing the relationship between the low-stage side compression ratio and high-stage compression ratio to total compression ratio to the presence of the injection parameters of the two-stage compressor 1 of displacement volume ratio 0.65. なお、図7と図8とには、全圧縮比に対する圧縮比の最適値も合わせて示す。 Incidentally, in FIGS. 7 and 8, is also shown the optimum value of the compression ratio to total compression ratio. なお、一般的に二段圧縮機1では、低段側圧縮比と高段側圧縮比とがそれぞれ全圧縮比の2乗根で、中間圧が二段圧縮機1の吸入圧力と吐出圧力との積の2乗根であるとき圧縮機効率が最もよいことが知られているので、低段側圧縮比と高段側圧縮比の最適値としてこの全圧縮比の2乗根を用いている。 In general two-stage compressor 1, with the square root of the low-stage side compression ratio and the high-stage compression ratio and each full compression ratio, the intermediate pressure and suction pressure of the two-stage compressor 1 and a discharge pressure since the square root compressor efficiency when a product is known to be the best uses a square root of the total compression ratio as the optimum value of the low-stage side compression ratio and the high-stage compression ratio .

排除容積比が0.85の二段圧縮機1の場合、図7に示すように、インジェクションを行うと、中間圧が上昇し、低段側圧縮比P d1 /P S1と高段側圧縮比P d2 /P S2がともに最適値に近づいて、圧縮機効率が改善できる。 If displacement volume ratio of the compressor 1 bunk 0.85, as shown in FIG. 7, when the injection, intermediate pressure increases, the low-stage side compression ratio P d1 / P S1 and the high-stage compression ratio approaching P d2 / P S2 are both optimum value, the compressor efficiency can be improved.
一方、排除容積比が0.65の二段圧縮機1の場合、インジェクションを行っても、低段側圧縮比と高段側圧縮比が変わらないので、圧縮機効率が改善できない。 On the other hand, if the displacement volume ratio of the compressor 1 bunk 0.65, even if the injection, since the low-stage side compression ratio and the high-stage compression ratio is not changed, the compressor efficiency can not be improved.

このような気液分離器を用いた二段圧縮インジェクションサイクルにおいては、特開2001−73976号公報に開示される排除容積比0.65は最適値ではない。 In such a gas-liquid separator two-stage compression injection cycle using, displacement volume ratio 0.65 as disclosed in JP 2001-73976 is not the optimum value. ガスインジェクションによる性能改善効果を得るためには、まず、第1の膨脹装置3で減圧された冷媒が飽和液圧P liq以下で、湿りガス状態になることが必要であり、また排除容積比に関する関係式(8)または関係式(11)が必要条件である。 To obtain the performance improvement effect by the gas injection, first, in a refrigerant decompressed by the first expansion device 3 is less than the saturation pressure P liq, it is necessary to become wet gas state, also relates to displacement volume ratio equation (8) or equation (11) is a prerequisite.

このようにすることにより、第1の膨張装置3で減圧された冷媒の圧力が飽和液圧以下となり、且つ、排除容積比が二段圧縮機1の吸入圧力の飽和液圧に対する比の等エントロピ指数の乗根以上となっているので、圧縮機効率と信頼性とを改善することができる。 In this way, the pressure of the reduced pressure refrigerant in the first expansion device 3 becomes saturated liquid pressure or under and, displacement volume ratio isentropic of ratio saturated liquid pressure of the suction pressure of the two-stage compressor 1 since a figure of root above, it is possible to improve the compressor efficiency and reliability.

また、中間圧が二段圧縮機1の吸入圧力と吐出圧力との積の2乗根より小さければ、二段圧縮機1の中間圧に冷媒を中間インジェクションすることによって、中間圧が二段圧縮機1の吸入圧力と吐出圧力との積の2乗根に漸近して圧縮機効率が改善されるとともに信頼性が向上する。 Further, if the intermediate pressure is less than the square root of the product of the suction pressure and the discharge pressure of the two-stage compressor 1, by which the intermediate injection refrigerant to an intermediate pressure of the two-stage compressor 1, the intermediate pressure is two-stage compression with the compressor efficiency by asymptotic to the square root of the product of the suction pressure and the discharge pressure of the machine 1 is improved and the reliability is improved. ここで、式(8)の飽和液圧P liqを二段圧縮機1の吸入圧力と吐出圧力との積の2乗根で置き換えれば、式(15)、式(16)が必要条件の式となる。 Here, by replacing the saturated liquid pressure P liq of formula (8) with the square root of the product of the suction pressure and the discharge pressure of the two-stage compressor 1, the formula (15), wherein the prerequisite formula (16) to become.

St2 /V St1 >(η V1 /η V2 )×(1+α×ρ d1 /ρ inj )×(P s1 /P d21/2n・・・(15) V St2 / V St1> (η V1 / η V2) × (1 + α × ρ d1 / ρ inj) × (P s1 / P d2) 1 / 2n ··· (15)
St2 /V St1 >(P s1 /P d21/2n・・・(16) V St2 / V St1> (P s1 / P d2) 1 / 2n ··· (16)

次に、実施の形態1に係わる中間圧に冷媒をインジェクションする効果について別の視点から説明する。 Next, a description from a different perspective on the effect of the injection refrigerant to an intermediate pressure according to the first embodiment.
表4には、二酸化炭素冷媒を用いた空調冷暖房設備の一般的な運転条件を示す。 Table 4 shows the general operating conditions of the air conditioning air conditioning equipment using a carbon dioxide refrigerant. 図9は、乾き度をパラメータとし、表4の冷房定格の条件の下、体積効率が1、中間加熱無しの理想条件で、二段圧縮機1の中間圧が最適値(P /P 1/2に一致する排除容積比を示すグラフである。 9, the degree of dryness as a parameter, under the condition of cooling rating of Table 4, the volumetric efficiency is 1, at ideal conditions without intermediate heating, intermediate pressure of the two-stage compressor 1 is the optimum value (P S / P d ) is a graph showing the displacement volume ratio matching 1/2. 図10は、乾き度をパラメータとし、表4の冷房定格の条件の下、体積効率が実体積効率、中間加熱の影響を考慮した実条件で、二段圧縮機1の中間圧が最適値(P /P 1/2に一致する排除容積比を示すグラフである。 Figure 10 is a dryness fraction as a parameter, under the condition of cooling rating of Table 4, the volumetric efficiency is the actual volumetric efficiency, the real conditions in consideration of the influence of the intermediate heating, intermediate pressure of the two-stage compressor 1 is the optimum value ( it is a graph showing the displacement volume ratio that matches P S / P d) 1/2.

冷媒インジェクション量の割合(α)が0%のとき、逆算した排除容積比は0.79となり、冷媒インジェクション量の割合が増加するほど最適な排除容積比は大きくなる。 When the proportion of the refrigerant injection amount (alpha) of 0%, the back-calculated displacement volume ratio 0.79, as the optimal displacement volume ratio proportion of the refrigerant injection amount increases the greater. 冷媒インジェクション量の割合が60%までインジェクション量を増やすと、排除容積比は理想条件で0.79〜1.12の範囲、実条件で0.84〜1.15の範囲で変化すると予測される。 When the ratio of the refrigerant injection amount increase injection amount up to 60% displacement volume ratio is expected to vary from the range of 0.79 to 1.12 in an ideal condition, the real conditions from 0.84 to 1.15 . 低段側吸入過熱度(SH)が大きいほど排除容積比は小さくなる、また、中間インジェクション冷媒の乾き度が小さいほど排除容積比は小さくなる傾向を示す。 As displacement volume ratio lower-stage suction superheat (SH) is large it becomes small and the higher displacement volume ratio dryness of the intermediate injection refrigerant is small showing a smaller tendency.

ガス主成分で冷媒をインジェクションする割合は冷凍サイクルの安定性を考慮すると少なすぎず大きすぎないことが必要であり、通常10%〜50%の範囲である。 Rate of injection of refrigerant gas mainly composed is required to be not too large nor too small in consideration of the stability of the refrigeration cycle, which is usually in the range of 10% to 50%. 最適な排除容積比はαによって変化するが、(P /P 1/2n以上の範囲が必要条件であり、通常考えられる冷房定格条件では0.8以上の値である。 Optimal displacement volume ratio varies by α, (P S / P d ) is a necessary condition 1 / 2n or more ranges, is 0.8 or more values in the cooling rated conditions which are normally considered.
また、このとき(P /P crt1/2n <(P /P 1/2nの関係式が成り立つので、気液分離インジェクションが成り立つための必要条件式(16)を満足する。 Moreover, since this time (P S / P crt) 1 / 2n <(P S / P d) 1 / 2n of the equation is satisfied, satisfying requirement equation for gas-liquid separation injection is established (16). 以上のように排除容積比を適切に設定すれば、中間インジェクションによって中間圧が最適値に近づいて圧縮機効率が改善する効果が得られる。 By appropriately setting the displacement volume ratio as described above, the effect of the compressor efficiency intermediate pressure by the intermediate injection approaches the optimum value is improved is obtained.

実施の形態2. The second embodiment.
図11は、この発明の実施の形態2に係わる気液分離器を用いた二段圧縮二段膨脹方式の超臨界蒸気圧縮式冷凍サイクルの回路図である。 Figure 11 is a circuit diagram of a supercritical vapor compression refrigeration cycle of the two-stage compression two-stage expansion method using a gas-liquid separator according to a second embodiment of the present invention.
この発明の実施の形態2に係わる超臨界蒸気圧縮式冷凍サイクルは、実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルに中間連結回路15に介設される逆止弁19と低段側から吐出する冷媒ガスを気液分離器4に戻す戻し回路18とを追加することが異なっており、それ以外は同様であるので、同様な部分に同じ符号を付記して説明は省略する。 Supercritical vapor compression refrigeration cycle according to the second embodiment of the invention, a check valve 19 which is interposed an intermediate connecting circuit 15 to the supercritical vapor compression refrigeration cycle according to the first embodiment from the lower stage refrigerant gas discharges are different to add a return circuit 18 back to the gas-liquid separator 4, since otherwise the same, described are indicated by the same reference numerals to like parts will be omitted.
実施の形態2に係わる超臨界蒸気圧縮式冷凍サイクルは、気液分離器4が液相状態となりインジェクションできないときのために、低段側回転圧縮要素11から吐出する冷媒ガスを中間連結回路15から分岐する逆止弁19と、分岐された冷媒ガスを気液分離器4に戻す戻り回路18を備える。 Supercritical vapor compression refrigeration cycle according to the second embodiment, in case the gas-liquid separator 4 can not be injection becomes liquid state, the refrigerant gas discharged from the low stage side rotary compressing element 11 from the intermediate connecting circuit 15 It comprises a check valve 19 for branching the return circuit 18 for returning the branched refrigerant gas in the gas-liquid separator 4.

このように低段側回転圧縮要素11から吐出する冷媒ガスを高段側吸入口S に送らずに、戻り回路18を経由して気液分離器4内に戻すと、気液分離器4内で液相冷媒と混合・熱交換して乾き度を高めて、気相の冷媒をインジェクション回路5から高段側吸入口S に注入することができる。 The refrigerant gas discharging this way from the low-stage side rotary compressing element 11 without transmitting the high-stage suction port S 2, when via the return circuit 18 back to the gas-liquid separator 4, the gas-liquid separator 4 to increase the degree of dryness by mixing and heat exchange with the liquid phase refrigerant in the inner, it is possible to inject the gas-phase refrigerant from the injection circuit 5 to the high-stage suction port S 2.
このようにすることにより、二段圧縮機1に中間インジェクションする冷媒主成分を気相状態に保つことができ、液相冷媒を多量注入することにより起る圧縮機効率の低下と信頼性の低下を防ぐことができる。 By doing so, I am possible to keep the refrigerant main component intermediate injection to two-stage compressor 1 to the gas-phase state, decrease in the decreased reliability of the compressor efficiency caused by large amounts injecting liquid refrigerant it is possible to prevent.

実施の形態3. Embodiment 3.
図12は、この発明の実施の形態3に係わる内部熱回収型熱交換器を用いた二段圧縮一段膨脹方式の超臨界蒸気圧縮式冷凍サイクルの回路図である。 Figure 12 is a circuit diagram of a supercritical vapor compression refrigeration cycle of the two-stage compression one-stage expansion method using an internal heat recovery type heat exchanger according to a third embodiment of the present invention.
この発明の実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルは、実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルと気液分離器4の代わりに内部熱回収型熱交換器10を用いることと、高圧放熱器2から第2の膨張装置7により低段側回転圧縮要素11への吸入圧力まで減圧することが異なっており、それ以外は同様であるので、同様な部分に同じ符号を付記して説明は省略する。 Supercritical vapor compression refrigeration cycle according to the third embodiment of the invention, using an internal heat recovery type heat exchanger 10 instead of the supercritical vapor compression refrigeration cycle and a gas-liquid separator 4 according to Embodiment 1 it and, by a second expansion device 7 from the high-pressure radiator 2 are different in a reduced pressure to the suction pressure to the low stage side rotary compressing element 11, since otherwise the same, the same reference numerals to like parts described with appended it will be omitted.

また、実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルは、インジェクション回路5と中間連結回路15の合流点に冷媒混合器14を備える。 Furthermore, the supercritical vapor compression refrigeration cycle according to the third embodiment includes a refrigerant mixer 14 at the confluence of the injection circuit 5 and the intermediate connecting circuit 15. 冷媒混合器14では、インジェクション回路5から供給される気相が主成分の冷媒を低段側回転圧縮要素11から吐出される冷媒に混合させてガス化してから、高段側回転圧縮要素12に吸入させる。 In the refrigerant mixture 14, the gas phase supplied from the injection circuit 5 is a refrigerant of the main component is mixed with from gasified refrigerant discharged from the low stage side rotary compressing element 11, the high stage side rotary compressing element 12 It is inhaled.

図13は、冷媒混合器14の一例の構成図である。 13 is a configuration diagram of an example of a refrigerant mixer 14.
冷媒混合器14は、図13に示すように、中間連結回路15の周囲に巻き付けられたねじり管方式の冷媒−冷媒熱交換器14aを備える。 Refrigerant mixer 14, as shown in FIG. 13, the refrigerant of the torsion tube type wrapped around the intermediate coupling circuit 15 - including a refrigerant heat exchanger 14a. この冷媒−冷媒熱交換器14aの一端はインジェクション回路5に接続され、他端が中間連結回路15の途中に接続される。 The refrigerant - one end of the refrigerant heat exchanger 14a is connected to the injection circuit 5 and the other end connected to the middle of the intermediate connecting circuit 15.
この冷媒混合器14では、インジェクション回路5から送られる液混じりのインジェクション冷媒と低段側回転圧縮要素11から吐出されたガス冷媒の間で熱交換され、乾き度が高められたインジェクション冷媒が中間連結回路15に合流され、高段側回転圧縮要素12に送られる。 In the refrigerant mixture 14, by heat exchange between the gas refrigerant discharged from the injection refrigerant and the low stage side rotary compressing element 11 of the liquid mingled sent from the injection circuit 5, the injection refrigerant is intermediate connection to dryness degree is increased is merged into the circuit 15, it is sent to the high stage side rotary compressing element 12.

図14は、冷媒混合器14の他の例の構成図である。 Figure 14 is a block diagram of another example of the refrigerant mixer 14.
また、別の冷媒混合器14は、図14に示すように、バッファタンク14bを備え、バッファタンク14b内で注入する冷媒が旋回流になるように、インジェクション回路5をバッファタンク14bに接続する。 Another refrigerant mixer 14, as shown in FIG. 14, a buffer tank 14b, the refrigerant to be injected with the buffer tank 14b is such that the swirling flow, to connect the injection circuit 5 to the buffer tank 14b. このようにインジェクション回路5から注入される冷媒が旋回されて混合されるので、十分に混合される。 Since the refrigerant to be injected from the injection circuit 5 are mixed is swiveled, it is mixed thoroughly.
なお、この冷媒混合機能をシェル内部で構成して二段圧縮機1をコンパクト化することも可能である。 It is also possible to compact a two-stage compressor 1 to constitute a refrigerant mixing function internally shell.

このようにすることにより、インジェクション回路5から液相冷媒を多量に流入する運転状態においても高段側回転圧縮要素12に液注入することを回避できるので信頼性の低下を防ぐことができる。 In this way, it is possible to prevent a decrease in reliability since it avoids the liquid injected into the high stage side rotary compressing element 12 even in the operating state a large amount flowing liquid-phase refrigerant from the injection circuit 5.
また、実施の形態3においても、排除容積比を実施の形態1と同様に適切に設定することにより、実施の形態1と同様に、二段圧縮機1に中間インジェクションする冷媒主成分を気相状態に保つことができ、液相冷媒を多量注入することによって起こる圧縮機効率の低下と信頼性の低下を防ぐことができる。 Also in the third embodiment, by setting similarly appropriately displacement volume ratio in the first embodiment, as in the first embodiment, the gas-phase refrigerant mainly composed of intermediate injection to two-stage compressor 1 it can keep, it is possible to prevent a decrease in reliability and decrease of the compression efficiency caused by large amounts injecting liquid refrigerant.

次に、実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルの効果を説明する。 Next, the effect of the supercritical vapor compression refrigeration cycle according to a third embodiment will be described.
図15は、実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルにおいて、インジェクション量の割合に対する暖房成績係数の関係を示すグラフである。 15, in a supercritical vapor compression refrigeration cycle according to Embodiment 3 is a graph showing the relationship between the heating coefficient of performance with respect to the ratio of the injection amount.
計算方法は、技術文献3(畝崎史武、「冷凍空調機器におけるシミュレーション技術」、冷凍、日本冷凍空調学会、2003年7月、第78巻、第909号、p.573−578)のサイクルシミュレーションを用いて、表4の暖房定格の運転条件で計算した。 Calculation method, the technical literature 3 (Sesakishibu, "the refrigerating and air-conditioning simulation technology in the equipment", refrigeration, Japan Society of Refrigerating and Air Conditioning Engineers, July 2003, Vol. 78, No. 909, p.573-578) of cycle using the simulation were calculated in the operating conditions of the heating rated in Table 4. 冷媒インジェクション量の割合(α)を約20%以上50%以下の範囲で暖房成績係数(COP )が最も良くて6%〜7%改善できる。 Heating coefficient of performance in a range rate (alpha) of about 20% to 50% of the refrigerant injection amount (COP H) can best be improved by 6% to 7%. αが10%以下、60%以上で性能が低下する。 α is 10% or less, performance decreases at least 60%.
図16は、実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルにおいて、過熱度(SH)をパラメータとするインジェクション量の割合に対する排除容積比の関係を示すグラフである。 16, in a supercritical vapor compression refrigeration cycle according to Embodiment 3 is a graph showing the relationship between the displacement volume ratio proportion of injection amount and the degree of superheat (SH) parameters.
図16に示す排除容量比は、まず表4の暖房定格の条件の下、適切なインジェクション量の割合、乾き度、中間圧力をそれぞれ計算し、これを満たす排除容積比を理想的な条件(η V1 =η V2 、中間加熱無し)から逆算して求める。 Void volume ratio shown in FIG. 16, first under the condition of heating rated in Table 4, the proportion of a suitable injection amount, the dryness, the intermediate pressure calculated respectively, ideal conditions the displacement volume ratio satisfying this (eta V1 = eta V2, obtained by back calculation from no intermediate heating).
図16から分かるように、インジェクション量の割合が増加するほど排除容積比は大きくなる。 As can be seen from Figure 16, displacement volume ratio as the ratio of the injection amount increases the greater. 標準的な低段側吸入口S の過熱度は10℃程度であり、過熱度を小さくすると排除容積比は大きくなる。 Standard low-stage side superheat of the suction port S 1 is about 10 ° C., displacement volume ratio to reduce the degree of superheat is increased.
また、図15と図16からわかるように、高い暖房成績係数(COP )が得られるのはインジェクション量の割合が20%以上50%以下の範囲であり、これは排除容積比が0.8〜1.1の範囲に相当する。 Moreover, as can be seen from FIGS. 15 and 16, the high heating coefficient of performance (COP H) is obtained in the range ratio less than 50% 20% of the injection amount, which is displacement volume ratio 0.8 It corresponds to the range of to 1.1.
通常は、排除容積比が1以下の範囲であるが、多量の冷媒を中間にインジェクションする場合には、排除容積比が1以上の場合も有効である。 Normally, displacement volume ratio is in the range of 1 or less, in the case of injection of a large amount of the refrigerant in the intermediate is also effective when the displacement volume ratio of 1 or more.

実施の形態4. Embodiment 4.
図17は、この発明の実施の形態4に係わる二段圧縮機の回転圧縮要素の断面図である。 Figure 17 is a cross-sectional view of a rotary compression element of two-stage compressor according to a fourth embodiment of the present invention.
低段側回転圧縮要素11と高段側回転圧縮要素12は、図17に示すように、共用のクランク軸30の主軸30aの周りに、それぞれ、ピン軸30b、シリンダ31、ローラ32、ベーン33、ベーン支持バネ34から構成される。 Low stage side rotary compressing element 11 and the high stage side rotary compressing element 12, as shown in FIG. 17, around the main shaft 30a of the crankshaft 30 of the shared, each pin shaft 30b, the cylinder 31, roller 32, the vane 33 , composed of the vane support spring 34. 通常、これらの部品寸法は排除容積の値に合わせて決定される。 Usually, these part dimensions are determined in accordance with the value of the displacement volume.

そして、この発明においては排除容積比を1になるようにインジェクション量を調整するので、これら部品寸法を共通化し、製造コストを低減することができる。 Since it adjusts the injection amount so that the displacement volume ratio 1 In the present invention, it is possible to commonly these part dimensions, to reduce the manufacturing cost.
但し、体積流量は低段側のほうが高段側より大きいので、吸入口35と吐出口36との内径寸法は低段側回転圧縮要素11のほうが高段側回転圧縮要素12より大きくすることが好ましい。 However, since the volume flow towards the lower stage is larger than the high-pressure stage, the inner diameter of the suction port 35 and the discharge port 36 be towards the low stage side rotary compressing element 11 is greater than the high stage side rotary compressing element 12 preferable. また、ローラ32の隙間寸法も低段側回転圧縮要素11と高段側回転圧縮要素12では異なる。 The roller 32 gap size or different in the low stage side rotary compressing element 11 and the high stage side rotary compressing element 12 of the.

実施の形態5. Embodiment 5.
図18は、この発明の実施の形態5に係わる超臨界蒸気圧縮式冷凍サイクルの回路図である。 Figure 18 is a circuit diagram of a supercritical vapor compression refrigeration cycle according to the fifth embodiment of the present invention.
この発明の実施の形態5に係わる超臨界蒸気圧縮式冷凍サイクルは、実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルの気液分離器4と第2の膨張装置7との間に内部熱交換器6を追加したことが異なっており、それ以外は同様であるので、同様な部分に同じ符号を付記して説明は省略する。 Supercritical vapor compression refrigeration cycle according to the fifth embodiment of the invention, the internal heat between the gas-liquid separator 4 supercritical vapor compression refrigeration cycle according to the first embodiment and the second expansion device 7 it is different that have added exchanger 6, since otherwise the same, described are indicated by the same reference numerals to like parts will be omitted.

実施の形態5に係わる超臨界蒸気圧縮式冷凍サイクルでは、気液分離器4で分離された液相の冷媒が内部熱交換器6内で、蒸発器8で吸熱加熱され気相状態まで蒸発され第2の気液分離器9で気液分離された気相冷媒との間で熱交換されるので、成績係数が実施の形態1の場合より向上する。 In the supercritical vapor compression refrigeration cycle according to the fifth embodiment, the refrigerant of the gas-liquid separator 4 in the separated liquid phase in the internal heat exchanger 6, is evaporated to gas phase is absorbed heated by the evaporator 8 since the heat exchange between the gas-phase refrigerant separated into gas and liquid in the second gas-liquid separator 9, the coefficient of performance is improved than in the first embodiment.

なお、実施の形態1乃至5のいずれにおいて、ロータリ式二段圧縮機の各段(低段と高段それぞれの)排除容積を理論排除容積(理論押しのけ量)V St thとして、式(17)から求めている。 Incidentally, in any of the first to fifth embodiments, each stage (low-stage and high-stage respectively) theoretical displacement volume of the displacement volume (theoretical displacement amount) of the rotary type two-stage compressor as V St th, formula (17) seeking from. 但し、Rはシリンダ半径、rはローラ半径、Lはシリンダ長さである。 Here, R is the cylinder radius, r is the roller radius, L is the cylinder length.

St th =π(R −r )L ・・・(17) V St th = π (R 2 -r 2) L ··· (17)

通常は、V St1 >V St2であるので、低段側シリンダ長さ(L )>高段側シリンダ長さ(L )とするか、または、低段側ローラ半径(r )<高段側ローラ半径(r )とすると2種類の回転圧縮要素を用いて構成する。 Normally, since it is V St1> V St2, or a low-stage side cylinder length (L 1)> high-stage side cylinder length (L 2), or, the low-stage-side roller radius (r 1) <High When stage roller radius (r 2) is constructed using two kinds of rotary compression elements.
さらに、通常、吸入口がベーン位置と重ならいように、吸入位相を遅れさせて吸入角度(θ :図17参照、単位は度=degreeで表示)をつける。 Further, usually, the suction port Ni no matter overlapping the vane position, intake angle allowed delay the suction phase (theta S: see FIG. 17, the unit is in degrees = degree) put. 30度未満の吸入角度θ を加味した実排除容積は理論排除容量と比べて小さいが、その差は1%以内に収まる。 Actual displacement volume in consideration of the suction angle theta S of less than 30 degrees is smaller than the theoretical exclusion volume, the difference falls within 1%. 従って、実施の形態1乃至5において排除容積比の範囲を指定した数値には1%以内の誤差が含まれている。 Therefore, the number that specifies the range of displacement volume ratio in the first to fifth embodiments includes an error of less than 1%.
また、吸入角度θ が約50度遅れると、実排除容積は理論排除容量より2〜3%小さくなるので、吸入角度θ を遅らせることにより、排除容積比を0.8〜1.0の範囲で可変できる。 Further, when the suction angle theta S is delayed approximately 50 degrees, since the actual displacement volume is 2-3% smaller than the theoretical exclusion volume, by delaying the intake angle theta S, a displacement volume ratio of 0.8 to 1.0 It can be varied in the range.

実施の形態1乃至5において得られた計算結果から、排除容積比が0.8以上、1以下であれば、圧縮機効率の改善と信頼性の向上ができると予測した。 From the calculation results obtained in the first to fifth embodiments, displacement volume ratio is 0.8 or more, as long as it is 1 or less, the improvement of improvement and reliability of the compressor efficiency was estimated to be. そこで、実施の形態1乃至5の超臨界蒸気圧縮式サイクルを超臨界で動作する二酸化炭素冷媒の冷暖房空調設備に適用する。 Therefore, applying a supercritical vapor compression cycle of the first to fifth embodiments in HVAC carbon dioxide refrigerant that operates in a supercritical. 例えば、冷房時室内温度25℃(ETは15℃)、冷房時室外温度30℃(T expは35℃)、暖房時室内温度20℃(T expは25℃)、暖房時室外温度1.2℃(ETは10℃)を仮定すると、冷房時圧縮比は約1.8〜1.9、暖房時圧縮比は約1.9〜2.1程度である。 For example, cooling at room temperature 25 ° C. (ET is 15 ° C.), cooling at outdoor temperature 30 ° C. (T exp is 35 ° C.), the heating time of the room temperature 20 ° C. (T exp is 25 ° C.), the heating time of outdoor temperature 1.2 ° C. (ET is 10 ° C.) assuming, cooling during compression ratio of about 1.8-1.9, the heating time of the compression ratio is about 1.9 to 2.1. このように、二酸化炭素冷媒はフロン冷媒に比べて比較的低い圧縮比条件で動作するため、排除容積比をフロン冷媒に比べて大きめに設定することが可能である。 Thus, carbon dioxide refrigerant is to operate at relatively low compression ratio condition in comparison with CFC refrigerants, it is possible to set larger than the displacement volume ratio Freon refrigerant. そして、二酸化炭素冷媒を低圧縮比の条件下で運転する場合、設計条件として排除容積比を0.8以上、1以下の範囲から選定すれば、圧縮機効率の改善と信頼性の向上が図られる。 When operating the carbon dioxide refrigerant under conditions of low compression ratio, 0.8 or more displacement volume ratio as design conditions, be selected from 1 or less in the range, improvement in the improvement and reliability of the compressor efficiency Figure It is.

また、実施の形態1乃至5の超臨界蒸気圧縮式サイクルを超臨界で動作する二酸化炭素冷媒のヒートポンプ給湯機に適用する。 Moreover, applying a supercritical vapor compression cycle of the first to fifth embodiments in the heat pump water heater of the carbon dioxide refrigerant that operates in a supercritical. 日本冷凍空調工業会標準規格の給湯定格の条件では、圧縮比は2.5程度である。 Under the conditions of the hot water supply rating of the Japan Refrigeration and Air Conditioning Industry Association standards, the compression ratio is about 2.5. このように、二酸化炭素冷媒はフロン冷媒に比べて比較的低い圧縮比条件で動作するため、排除容積比をフロン冷媒に比べて大きめに設定することが可能である。 Thus, carbon dioxide refrigerant is to operate at relatively low compression ratio condition in comparison with CFC refrigerants, it is possible to set larger than the displacement volume ratio Freon refrigerant.

また、実施の形態1乃至5において得られた計算結果と実験結果は、高圧シェル型圧縮機を想定して説明したが、低圧シェル型や中間圧シェル型においても同様の計算結果と事件結果が得られ、同様な方法で排除容積比を設定すれば、中間圧が適切に設定され、圧縮機効率が高く且つ信頼性に優れた冷暖房空調設備やヒートポンプ給湯機を提供することができる。 The calculation results and experimental results obtained in the first to fifth embodiments have been described assuming a high pressure shell type compressor, the same calculation results incident results in a low pressure shell type and intermediate pressure shell the resulting, by setting the displacement volume ratio in a similar way, the intermediate pressure is set appropriately, it is possible to provide a compressor efficiency is high and excellent reliability HVAC or heat pump water heater.

また、実施の形態1乃至5において得られた計算結果と実験結果は、ロータリ式二段圧縮機を想定して説明したが、スクロール式、スイング式、レシプロ式、その他圧縮方式の場合も、ロータリ式と同様の計算結果と実験結果が得られ、同様な方法で排除容積比を設定すれば、中間圧が適切に設定され、圧縮機効率が高く且つ信頼性に優れた冷暖房空調設備やヒートポンプ給湯機を提供することができる。 The obtained calculation results and experimental results in the first to fifth embodiments have been described assuming a rotary two-stage compressor, a scroll type, swing type, reciprocating, even if other compression method, rotary similar to the formula calculation results and experimental results obtained, by setting the displacement volume ratio in the same manner, is set an intermediate pressure is suitably, the compressor efficiency and excellent high and reliable HVAC or heat pump water it is possible to provide a machine.

また、実施の形態1乃至5においては、高圧側が超臨界で動作するサイクルの冷媒として二酸化炭素を用いたが、臨界温度の特性から同様に超臨界サイクルを構成しうる冷媒としてはエタン、アセチレン、二酸化窒素、クロロトリフルオロメタン(R13)、トリフルオロメタン(R23)、フルオロメタン(R41)を用いることができる。 In the first to fifth embodiments, although the high pressure side is used carbon dioxide as a refrigerant cycle which operates at supercritical ethane as a refrigerant capable of constituting a supercritical cycle in the same manner from the characteristics of the critical temperature, acetylene, nitrogen dioxide, chlorotrifluoromethane (R13), trifluoromethane (R23), may be used trifluoromethane (R41).
また、二酸化炭素にジフオロメタン(R32)を20w%以内で混入した混合冷媒も用いることができる。 Also, mixed refrigerant mixed carbon dioxide Jifuorometan the (R32) within 20 w% can also be used. これらの冷媒を用いた超臨界蒸気圧縮式冷凍サイクルにおいても、実施の形態1乃至5と同様な方法で排除容積比を設定すれば、中間圧が適切に設定され、圧縮機効率が高く且つ信頼性に優れる。 Also in the supercritical vapor compression refrigeration cycle using these refrigerants, by setting the displacement volume ratio in the same manner as in Embodiments 1 to 5 the method, the intermediate pressure is properly set, the compressor efficiency is high and reliable excellent sex. 但し、排除容積比の数値範囲は冷媒により多少異なる。 However, the numerical range of displacement volume ratio somewhat with the refrigerant.

この発明の実施の形態1に係わる気液分離器を用いた超臨界蒸気圧縮式冷凍サイクルの回路図である。 It is a circuit diagram of a supercritical vapor compression refrigeration cycle using a gas-liquid separator according to the first embodiment of the present invention. この発明の実施の形態1に係わる超臨界蒸気圧縮式冷凍サイクルのP−h線図である。 It is a P-h diagram of a supercritical vapor compression refrigeration cycle according to the first embodiment of the present invention. この発明の実施の形態1に係わる二段圧縮機の高段側回転圧縮要素の吸入口でのP−h線図である。 It is a P-h diagram at the intake of the high-stage side rotary compressing element of the two-stage compressor according to the first embodiment of the present invention. 排除容積比およびインジェクションの有無をパラメータとする全圧縮比に対する圧縮機効率の関係を示すグラフである。 Is a graph showing the relationship between the compressor efficiency for all compression ratio whether the displacement volume ratio and injection as parameters. 排除容積比およびインジェクションの有無をパラメータとする全圧縮比に対する給湯加熱成績係数の関係を示すグラフである。 Is a graph showing the relationship between the hot water heating coefficient of performance whether the displacement volume ratio and injection to total compression ratio of a parameter. 排除容積比をパラメータとする全圧縮比に対するガスインジェクション量の割合の関係を示すグラフである。 Is a graph showing the relationship between the proportion of the gas injection amount to the total compression ratio of a displacement volume ratio as a parameter. 排除容積比0.85の二段圧縮機におけるインジェクションの有無をパラメータとする全圧縮比に対する低段側圧縮比および高段側圧縮比の関係を示すグラフである。 Is a graph showing the relationship between the low-stage side compression ratio and high-stage compression ratio whether the injection of the two-stage compressor to the total compression ratio of a parameter of the displacement volume ratio 0.85. 排除容積比0.65の二段圧縮機におけるインジェクションの有無をパラメータとする全圧縮比に対する低段側圧縮比および高段側圧縮比の関係を示すグラフである。 Is a graph showing the relationship between the low-stage side compression ratio and high-stage compression ratio to total compression ratio of a parameter whether the injection of the two-stage compressor of the displacement volume ratio 0.65. 乾き度をパラメータとし、理想条件で二段圧縮機の中間圧が最適値に一致する排除容積比を示すグラフである。 The dryness fraction as a parameter, a graph showing the displacement volume ratio intermediate pressure of the two-stage compressor in the ideal condition matches the optimum value. 乾き度をパラメータとし、実条件で二段圧縮機の中間圧が最適値に一致する排除容積比を示すグラフである。 The dryness fraction as a parameter, a graph showing the displacement volume ratio intermediate pressure of the two-stage compressor in real conditions is equal to the optimum value. この発明の実施の形態2に係わる気液分離器を用いた二段圧縮二段膨脹方式の超臨界蒸気圧縮式冷凍サイクルの回路図である。 It is a circuit diagram of a supercritical vapor compression refrigeration cycle of the two-stage compression two-stage expansion method using a gas-liquid separator according to a second embodiment of the present invention. この発明の実施の形態3に係わる内部熱回収型熱交換器を用いた二段圧縮一段膨脹方式の超臨界蒸気圧縮式冷凍サイクルの回路図である。 It is a circuit diagram of a supercritical vapor compression refrigeration cycle of the two-stage compression one-stage expansion method using an internal heat recovery type heat exchanger according to a third embodiment of the present invention. 冷媒混合器の一例の構成図である。 It is a configuration diagram of an example of a refrigerant mixer. 冷媒混合器の他の例の構成図である。 It is a block diagram of another example of the refrigerant mixer. 実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルにおいて、インジェクション量の割合に対する暖房成績係数の関係を示すグラフである。 Supercritical vapor compression refrigeration cycle according to Embodiment 3 is a graph showing the relationship between the heating coefficient of performance with respect to the ratio of the injection amount. 実施の形態3に係わる超臨界蒸気圧縮式冷凍サイクルにおいて、過熱度をパラメータとするインジェクション量の割合に対する排除容積比の関係を示すグラフである。 Supercritical vapor compression refrigeration cycle according to Embodiment 3 is a graph showing the relationship between the displacement volume ratio proportion of injection amount of superheat parameters. この発明の実施の形態4に係わる二段圧縮機の回転圧縮要素の断面図である。 It is a cross-sectional view of a rotary compression element of two-stage compressor according to a fourth embodiment of the present invention. この発明の実施の形態5に係わる超臨界蒸気圧縮式冷凍サイクルの回路図である。 It is a circuit diagram of a supercritical vapor compression refrigeration cycle according to the fifth embodiment of the present invention.

符号の説明 DESCRIPTION OF SYMBOLS

1 二段圧縮機、2 高圧放熱器、3、7 膨張装置、4、9 気液分離器、5 インジェクション回路、6 内部熱交換器、8 蒸発器、10 内部熱回収型熱交換器、11 低段側回転圧縮要素、12 高段側回転圧縮要素、13 密閉容器、14 冷媒混合器、14a 冷媒−冷媒熱交換器、14b バッファタンク、15 中間連結回路、16 流量調整弁、18 戻り回路、19 逆止弁、20、21、22、24、25 温度計、23 液相側配管、30a 主軸、30b ピン軸、31 シリンダ、32 ローラ、33 ベーン、34 ベーン支持バネ、35 吸入口、36 吐出口。 1 two-stage compressor, 2 a high pressure radiator, 3,7 expansion device, 4,9 gas-liquid separator, 5 injection circuit, 6 an internal heat exchanger, 8 evaporator, 10 internal heat recovery type heat exchanger, 11 low stage side rotary compressing element, 12 the high stage side rotary compressing element 13 closed vessel 14 coolant mixer, 14a refrigerant - refrigerant heat exchanger, 14b buffer tank, 15 an intermediate connecting circuit, 16 a flow rate control valve, 18 return circuit, 19 check valves, 20,21,22,24,25 thermometer, 23 liquid phase pipe, 30a spindle, 30b pin shaft, 31 cylinder, 32 roller, 33 vanes, 34 vane support spring, 35 suction port, 36 discharge port .

Claims (8)

  1. 低圧の冷媒が低段側回転圧縮要素で中間圧まで圧縮された冷媒が中間連結回路を経由して高段側回転圧縮要素に吸入され、上記高段側回転圧縮要素で高圧まで圧縮される二段圧縮機と、高圧に圧縮される冷媒を飽和液圧以下まで減圧する第1の膨張装置と、飽和液圧以下で湿りガス状態になった冷媒を気液分離する気液分離器と、気液分離後の気相側冷媒を上記中間連結回路にインジェクションするインジェクション回路と、気液分離後の液相側冷媒を低圧まで減圧する第2の膨張装置と、低圧まで減圧される冷媒を蒸発させる蒸発器と、を有する超臨界蒸気圧縮式冷凍サイクルにおいて、 Two low-pressure refrigerant refrigerant compressed by the low stage side rotary compressing element to an intermediate pressure is sucked into the high stage side rotary compressing element through the intermediate connecting circuit, it is compressed to a high pressure in the high stage side rotary compressing element and stage compressor, first and expansion device for reducing the pressure of the refrigerant compressed to a high pressure to under saturated liquid pressure or a gas-liquid separator to a refrigerant liquid separation became wet gas state under saturated liquid pressure or, air and injection circuit for injecting a gas-phase refrigerant after liquid separation to the intermediate connecting circuit, a second expansion device for decompressing the liquid-phase refrigerant after gas-liquid separation to a low pressure to evaporate the refrigerant to be reduced to a low pressure an evaporator in supercritical vapor compression refrigeration cycle having,
    上記高段側回転圧縮要素の排除容積の上記低段側回転圧縮要素の排除容積に対する排除容積比は、上記二段圧縮機の吸入圧力の上記第1の膨張装置における冷媒飽和液圧に対する比の等エントロピ指数乗根以上であり、 Displacement volume ratio displacement volume of the low stage side rotary compressing element of the displacement volume of the high stage side rotary compressing element, the ratio of refrigerant saturated liquid pressure in the first expansion device of the suction pressure of the two-stage compressor is the isentropic index root or more,
    上記第1の膨張装置または上記第2の膨張装置の開度または吸入加熱度を、上記中間連結回路にインジェクションする冷媒を気相状態に保つように制御することを特徴とする超臨界蒸気圧縮式冷凍サイクル。 The opening or suction superheat of the first expansion device or the second expansion device, a supercritical vapor compression, characterized by controlling so as to maintain the refrigerant is injected to the intermediate connecting circuit to the gas-phase state refrigeration cycle.
  2. 低圧の冷媒が低段側回転圧縮要素で中間圧まで圧縮される冷媒が中間連結回路を経由して高段側回転圧縮要素に吸入され、上記高段側回転圧縮要素で高圧まで圧縮される二段圧縮機と、高圧に圧縮された冷媒を冷却して2系統に分割する高圧放熱器と、分割された一方の冷媒を中間圧まで減圧する第1の膨張装置と、上記第1の膨張装置で中間圧まで減圧された冷媒が内部熱回収する熱交換器と、上記内部熱回収した冷媒を上記中間連結回路にインジェクションするインジェクション回路と、分割された他方の冷媒を低圧まで減圧する第2の膨張装置と、低圧まで減圧された冷媒を蒸発させる蒸発器と、を有する超臨界蒸気圧縮式冷凍サイクルにおいて、 Two low-pressure refrigerant refrigerant compressed by the low stage side rotary compressing element to an intermediate pressure is sucked by way of the intermediate connecting circuit high stage side rotary compressing element is compressed to a high pressure in the high stage side rotary compressing element and stage compressor, a high pressure radiator for dividing the refrigerant compressed to a high pressure to the two systems cooling, a first expansion device for reducing the pressure divided one refrigerant to an intermediate pressure, said first expansion device in a heat exchanger refrigerant depressurized to an intermediate pressure is internal heat recovery, the injection circuit for injecting the refrigerant recovered the internal heat to the intermediate connecting circuit, a second depressurizing the other split portion of the refrigerant to a low pressure an expansion device, an evaporator for evaporating the decompressed refrigerant to a low pressure, the supercritical vapor compression refrigeration cycle having,
    上記高段側回転圧縮要素の排除容積の上記低段側回転圧縮要素の排除容積に対する排除容積比は、上記二段圧縮機の吸入圧力の上記第1の膨張装置における冷媒飽和液圧に対する比の等エントロピ指数乗根以上であり、 Displacement volume ratio displacement volume of the low stage side rotary compressing element of the displacement volume of the high stage side rotary compressing element, the ratio of refrigerant saturated liquid pressure in the first expansion device of the suction pressure of the two-stage compressor is the isentropic index root or more,
    上記第1の膨張装置の開度または吸入加熱度を、上記中間連結回路にインジェクションする冷媒の乾き度を高い状態に保つように制御することを特徴とする超臨界蒸気圧縮式冷凍サイクル。 The opening degree or inhalation heating of the first expansion device, a supercritical vapor compression refrigeration cycle and controls to keep the dryness of the refrigerant is injected to the intermediate connecting circuit to a high state.
  3. 低圧の冷媒が低段側回転圧縮要素で中間圧まで圧縮される冷媒が中間連結回路を経由して高段側回転圧縮要素に吸入され、上記高段側回転圧縮要素で高圧まで圧縮される二段圧縮機と、高圧に圧縮される冷媒を飽和液圧以下まで減圧する第1の膨張装置と、飽和液圧以下で湿りガス状態になった冷媒を気液分離する気液分離器と、気液分離後の気相側冷媒を上記中間連結回路にインジェクションするインジェクション回路と、上記気液分離後の液相側冷媒を低圧まで減圧する第2の膨張装置と、低圧まで減圧される冷媒を蒸発させる蒸発器と、を有する超臨界蒸気圧縮式冷凍サイクルにおいて、 Two low-pressure refrigerant refrigerant compressed by the low stage side rotary compressing element to an intermediate pressure is sucked by way of the intermediate connecting circuit high stage side rotary compressing element is compressed to a high pressure in the high stage side rotary compressing element and stage compressor, first and expansion device for reducing the pressure of the refrigerant compressed to a high pressure to under saturated liquid pressure or a gas-liquid separator to a refrigerant liquid separation became wet gas state under saturated liquid pressure or, air evaporated with injection circuit for injecting a gas-phase refrigerant after liquid separation to the intermediate connecting circuit, a second expansion device for decompressing the liquid-phase refrigerant after the gas-liquid separator to a low pressure, the refrigerant is depressurized to a low pressure an evaporator for, in a supercritical vapor compression refrigeration cycle having,
    上記高段側回転圧縮要素の排除容積の上記低段側回転圧縮要素の排除容積に対する排除容積比は、上記二段圧縮機の吸入圧力の上記二段圧縮機の吐出圧力に対する比の等エントロピ指数の2倍の乗根以上であり、 Displacement volume ratio displacement volume of the low stage side rotary compressing element of the displacement volume of the high stage side rotary compressing element, isentropic exponent ratio discharge pressure of the two-stage compressor suction pressure of the two-stage compressor is 2 times the root or more,
    上記第1の膨張装置または上記第2の膨張装置の開度を、上記中間連結回路にインジェクションする冷媒の圧力を上記吸入圧力と上記吐出圧力との積の2乗根に近づくように制御することを特徴とする超臨界蒸気圧縮式冷凍サイクル。 The opening of the first expansion device or the second expansion device, be controlled so as to approach the pressure of the refrigerant is injected to the intermediate connecting circuit to the square root of the product of the suction pressure and the discharge pressure supercritical vapor compression refrigeration cycle, wherein.
  4. 低圧の冷媒が低段側回転圧縮要素で中間圧まで圧縮される冷媒が中間連結回路を経由して高段側回転圧縮要素に吸入され、上記高段側回転圧縮要素で高圧まで圧縮される二段圧縮機と、高圧に圧縮された冷媒を冷却して2系統に分割する高圧放熱器と、分割された一方の冷媒を中間圧まで減圧する第1の膨張装置と、上記第1の膨張装置で中間圧まで減圧された冷媒に上記高圧放熱器から出力される冷媒から熱交換して内部熱回収する熱交換器と、上記内部熱回収した冷媒を上記中間連結回路にインジェクションするインジェクション回路と、分割された他方の冷媒を低圧まで減圧する第2の膨張装置と、低圧まで減圧された冷媒を蒸発させる蒸発器と、を有する超臨界蒸気圧縮式冷凍サイクルにおいて、 Two low-pressure refrigerant refrigerant compressed by the low stage side rotary compressing element to an intermediate pressure is sucked by way of the intermediate connecting circuit high stage side rotary compressing element is compressed to a high pressure in the high stage side rotary compressing element and stage compressor, a high pressure radiator for dividing the refrigerant compressed to a high pressure to the two systems cooling, a first expansion device for reducing the pressure divided one refrigerant to an intermediate pressure, said first expansion device in a heat exchanger for internal heat recovery from the refrigerant output from the high-pressure radiator to decompressed refrigerant to an intermediate pressure by the heat exchange, the injection circuit for injecting the refrigerant recovered the internal heat to the intermediate connecting circuit, a second expansion device for decompressing the other split portion of the refrigerant to a low pressure, an evaporator for evaporating the decompressed refrigerant to a low pressure, the supercritical vapor compression refrigeration cycle having,
    上記高段側回転圧縮要素の排除容積の上記低段側回転圧縮要素の排除容積に対する排除容積比は、上記二段圧縮機の吸入圧力の上記二段圧縮機の吐出圧力に対する比の等エントロピ指数の2倍の乗根以上であり、 Displacement volume ratio displacement volume of the low stage side rotary compressing element of the displacement volume of the high stage side rotary compressing element, isentropic exponent ratio discharge pressure of the two-stage compressor suction pressure of the two-stage compressor is 2 times the root or more,
    上記第1の膨張装置の開度を、上記中間連結回路にインジェクションする冷媒の圧力を上記吸入圧力と上記吐出圧力との積の2乗根に近づくように制御することを特徴とする超臨界蒸気圧縮式冷凍サイクル。 Supercritical steam and controls so as to approach the opening of the first expansion device, the pressure of the refrigerant is injected to the intermediate connecting circuit to the square root of the product of the suction pressure and the discharge pressure compression refrigeration cycle.
  5. 冷媒は、二酸化炭素からなり、 Refrigerant, made from carbon dioxide,
    上記排除容積比が0.8以上、1.1以下であることを特徴とする請求項1乃至4のいずれか一項に記載する超臨界蒸気圧縮式冷凍サイクル。 The displacement volume ratio is 0.8 or higher, supercritical vapor compression refrigeration cycle according to any one of claims 1 to 4, characterized in that more than 1.1.
  6. 上記排除容積比が1であることを特徴とする請求項1乃至4のいずれか一項に記載する超臨界蒸気圧縮式冷凍サイクル。 Supercritical vapor compression refrigeration cycle according to any one of claims 1 to 4, characterized in that the displacement volume ratio is 1.
  7. 冷媒が二酸化炭素からなる請求項1乃至6のいずれか一項に記載する超臨界蒸気圧縮式冷凍サイクルを用いることを特徴とする冷暖房空調設備。 HVAC, which comprises using a supercritical vapor compression refrigeration cycle according to any one of claims 1 to 6 refrigerant consisting of carbon dioxide.
  8. 冷媒が二酸化炭素からなる請求項1乃至6のいずれか一項に記載する超臨界蒸気圧縮式冷凍サイクルを用いることを特徴とするヒートポンプ式給湯機。 Heat pump water heater characterized by using a supercritical vapor compression refrigeration cycle according to any one of claims 1 to 6 refrigerant consisting of carbon dioxide.
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