JP3835141B2 - heat pump - Google Patents

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Publication number
JP3835141B2
JP3835141B2 JP2000264861A JP2000264861A JP3835141B2 JP 3835141 B2 JP3835141 B2 JP 3835141B2 JP 2000264861 A JP2000264861 A JP 2000264861A JP 2000264861 A JP2000264861 A JP 2000264861A JP 3835141 B2 JP3835141 B2 JP 3835141B2
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Prior art keywords
refrigerant
pressure
compressor
throttle valve
variable throttle
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Expired - Fee Related
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JP2000264861A
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JP2002081771A (en
Inventor
申也 野呂
久介 榊原
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Denso Corp
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Denso Corp
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant the refrigerant being carbon dioxide
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2309/00—Gas cycle refrigeration machines
    • F25B2309/06—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide
    • F25B2309/061—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide with cycle highest pressure above the supercritical pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2339/00—Details of evaporators; Details of condensers
    • F25B2339/04—Details of condensers
    • F25B2339/047—Water-cooled condensers
    • 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
    • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
    • F25B2341/001—Ejectors not being used as compression device
    • F25B2341/0012—Ejectors with the cooled primary flow at high pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
    • F25B2341/001—Ejectors not being used as compression device
    • F25B2341/0013—Ejector control arrangements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/17—Control issues by controlling the pressure of the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2513—Expansion valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/19—Pressures
    • F25B2700/193—Pressures of the compressor
    • F25B2700/1931—Discharge pressures
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2116—Temperatures of a condenser
    • F25B2700/21161—Temperatures of a condenser of the fluid heated by the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
    • F25B2700/21—Temperatures
    • F25B2700/2116—Temperatures of a condenser
    • F25B2700/21163—Temperatures of a condenser of the refrigerant at the outlet of the condenser

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Chemical Kinetics & Catalysis (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Heat-Pump Type And Storage Water Heaters (AREA)

Description

【０００１】
【発明の属する技術分野】
本発明は、例えば、給湯器の給湯用水を加熱する加熱手段として用いて好適な、超臨界ヒートポンプサイクルで作動するヒートポンプに関するものである。
【０００２】
【従来の技術】
従来より、例えば、給湯器の給湯用水を加熱するのに、高圧側の冷媒圧力を臨界圧力以上に加圧する超臨界ヒートポンプサイクルを用いたヒートポンプが知られている。
【０００３】
図６に示すように、ヒートポンプ１０は、圧縮機１１、水熱交換器１３、膨張弁２３、蒸発器１８、気液分離器１６を順次冷媒配管で接続したもので、水熱交換器１３の負荷流体として給湯用水を流入させ、高温高圧の冷媒と熱交換させることで給湯用水を加熱する。この水熱交換器１３で加熱された給湯用水はタンク２内に貯留され、使用時に温度調節した後、使用者に供給するようにしている。
【０００４】
そして、このヒートポンプ１０は、サイクル効率の高い領域でヒートポンプサイクルを運転するために、水熱交換器１３に流入する給湯用水と水熱交換器１３から流出する冷媒との温度差によってサイクルの高圧制御が行なわれている。
【０００５】
【発明が解決しようとする課題】
ところで、本発明者らは、このヒートポンプ１０のサイクル効率を更に向上させるために、エジェクタを用いて圧縮機１１への吸入圧力を上昇させ、圧縮機１１の圧縮仕事を低減させることを考案したが、この場合、圧縮機１１の吐出側の冷媒温度が低下するために、水熱交換器１３での熱交換能力が低下してしまうという問題があった。
【０００６】
本発明の目的は、上記問題に鑑み、ヒートポンプサイクルにおける圧縮機の圧縮仕事を低減し、且つ、水熱交換器の熱交換能力を確保できるヒートポンプを提供することにある。
【０００７】
【課題を解決するための手段】
本発明は上記目的を達成するために、以下の技術的手段を採用する。
【０００８】
請求項１に記載の発明では、気液分離器（１６）より吸入した気相冷媒を圧縮機（１１）で臨界圧力以上に加圧して、高圧側熱交換器（１３）により冷媒と負荷流体とを対向流として熱交換させるヒートポンプにおいて、高圧側熱交換器（１３）から流出する冷媒、および、気液分離器（１６）の液相冷媒を蒸発させる蒸発器（１８）から流出する冷媒を吸入し、圧縮機（１１）の中間圧力まで減圧した後、気液分離器（１６）に流入させるエジェクタ（１５）と、圧縮機（１１）の高圧側となるように高圧側熱交換器（１３）の下流側に設けられ、冷媒流量を可変する第１の可変絞り弁（１４）と、圧縮機（１１）への吸入圧力を調整するように前記圧縮機（１１）の低圧側に設けられ、冷媒流量を可変する第２の可変絞り弁（１７）とを設け、この第１、第２の可変絞り弁（１４、１７）の開度を制御する制御手段（２２）により、負荷流体との熱交換能力を確保するように圧縮機（１１）の高圧側の圧力を変化させるために第１の可変絞り弁（１４）の開度を制御し、吸入圧力を調整することで圧縮機（１１）の高圧側の圧力を制御するように第２の可変絞り弁（１７）の開度を制御するようにしたことを特徴としている。
【０００９】
これにより、圧縮機（１１）の吸入圧力がエジェクタ（１５）により上昇するので、圧縮仕事を低減できる。加えて、第1の可変絞り弁（１４）で圧縮機（１１）の吐出圧力を可変し、吐出冷媒の温度を制御できるので、負荷流体との熱交換能力を常に確保できる。
【００１０】
請求項２に記載の発明では、制御手段（２２）により、高圧側熱交換器（１３）に流入する負荷流体と高圧側熱交換器（１３）から流出する冷媒との温度差が目標温度差ΔＴとなるように、第１の可変絞り弁（１４）の開度を制御することを特徴としている。
【００１１】
これにより、目標温度差ΔＴが常に一定になるように制御するので、効率よく負荷流体を一定温度に昇温できる。
【００１２】
請求項３に記載の発明では、制御手段（２２）により、圧縮機（１１）の吐出圧力、あるいは吐出圧力に相関する物理量に応じて、第２の可変絞り弁（１７）の開度を制御することを特徴としている。
【００１３】
これにより、例えば、外気温度の変化に対して第２可変絞り弁で吸入圧力を可変させ吐出圧力を制御するので、吸入圧力が高い場合には、吸入圧力を下げることにより吐出圧力を下げて、ヒートポンプサイクル内の耐圧上の許容圧力を超えない様に保護できる。また、吸入圧力が低い場合には、吸入圧力を上げて圧縮仕事を低減できる。
【００１４】
請求項４に記載の発明では、第２の可変絞り弁（１７）は、蒸発器（１８）からエジェクタ（１５）に流入する冷媒流量を調整することを特徴としている。また、請求項５に記載の発明では、第２の可変絞り弁（１７）は、気液分離器（１６）の下流側に設けられていることを特徴としている。また、請求項６に記載の発明では、制御手段（２２）は、熱交換能力を確保するために冷媒温度がＴｒｉになるように第１の可変絞り弁（１４）の開度を制御することで圧縮機（１１）からの吐出圧力がＰ１より高いＰ２になると、第２の可変絞り弁（１７）の開度を小さくし、また、熱交換能力を確保するために冷媒温度がＴｒｉになるように第１の可変絞り弁（１４）の開度を制御することで圧縮機（１１）からの吐出圧力がＰ１より低いＰ３になると、第２の可変絞り弁（１７）の開度を大きくすることを特徴としている。尚、上記各手段の括弧内の符号は、後述する実施形態記載の具体的手段との対応関係を示すものである。
【００１５】
【発明の実施の形態】
（第１実施形態）
本発明の第１実施形態を図１に示す。第１実施形態は、本発明のヒートポンプ１０を熱源とする、ヒートポンプ式給湯器１に適用したものである。
【００１６】
ヒートポンプ式給湯器１は、加熱された給湯用水をタンク２内に貯留しておき、使用時にタンク２内から給湯用水を取出し、温度調節して使用者に供給するシステムである。
【００１７】
タンク２は、耐食性に優れた金属製（例えばステンレス製）で断熱構造を有し、高温の給湯用水を長時間に渡って保温することができる。尚、タンク２内に貯留される給湯用水は、キッチンや風呂等で直接使用してもよいが、給湯用以外に、例えば床暖房用、室内空調用等の熱源として利用することもできる。
【００１８】
電動ポンプ３は、タンク２とヒートポンプ１０の高圧側熱交換器（以下、水熱交換器と呼ぶ）１３とを環状に接続する温水配管４に設けられて、タンク２と水熱交換器１３との間で給湯用水を循環させるとともに、内蔵するモータの回転数に応じて循環水量を調節する。
【００１９】
ヒートポンプ１０は、給湯用水に対する熱源として冷媒温度（本実施形態では１３０℃）を充分に確保するために高圧側の冷媒圧力が臨界圧力以上となるようにしており、圧縮機１１、水熱交換器１３、エジェクタ１５、気液分離器１６、蒸発器１８等によって構成され、冷媒として臨界圧力の低い二酸化炭素（ＣＯ2）を使用している。
【００２０】
圧縮機１１は、内蔵する電動モータ（図示せず）によって駆動され、後述する気液分離器１６から吸入した気相冷媒を臨界圧力以上に圧縮して吐出する。
【００２１】
オイル分離器１２は、圧縮機１１と水熱交換器１３との間に設けられ、圧縮機１１から吐出された冷媒とオイルとを分離し、余剰オイルを蓄えるとともに順次オイルを圧縮機１１に戻し、分離された冷媒のみを水熱交換器１３に流入させる。
【００２２】
水熱交換器１３は、圧縮機１１より吐出された高圧の気相冷媒と負荷流体としての給湯用水とを熱交換するもので、冷媒の流れ方向と給湯用水の流れ方向とが対向するように構成されている。
【００２３】
そして、水熱交換器１３の冷媒流入側、流出側には、それぞれ圧縮機１１から吐出される冷媒の吐出圧力を検出する圧力センサ１９、水熱交換器１３に流入する給湯用水の温度を検出する水用温度センサ２０、水熱交換器１３より流出する冷媒の温度を検出する冷媒用温度センサ２１が設けられ、各センサ１９〜２１の検出信号が電子制御装置（以下、ＥＣＵと呼ぶ）２２に入力される。
【００２４】
ＥＣＵ２２は、ヒートポンプ１０を効率よく運転できるように、後述する第１、第２の可変絞り弁１４、１７の開度を制御する。
【００２５】
第１可変絞り弁１４は、圧縮機１１の高圧側となるように、水熱交換器１３の下流側に設けられ、ＥＣＵ２２の制御信号に基づいて弁の開度が可変制御され、圧縮機１１の吐出圧力を調整する。
【００２６】
更に、第１可変絞り弁１４の下流側には、ノズル１５ａ、吸引部１５ｂ、混合部１５ｃ、ディフューザ１５ｄとから成るエジェクタ１５が設けられている。エジェクタ１５は、水熱交換器１３で熱交換された冷媒を圧縮機１１の中間圧力まで減圧させるものである。まず、水熱交換器１３から流出する冷媒が滑らかに収縮するノズル１５ａにより絞り損失を伴わずに減圧加速される。次に、ノズル１５ａで加速された冷媒流れの吸引効果により、後述する蒸発器１８から流出する冷媒が、吸引部１５ｂから吸引され、混合部１５ｃで混合し圧力を上昇させ、滑らかに拡大するディフューザ１５ｄにより減速され、中間圧力まで上昇する。
【００２７】
気液分離器１６は、エジェクタ１５から流出する冷媒を気液分離してサイクル内の余剰冷媒を蓄えるとともに、気相冷媒を圧縮機１１に吸入させ、また、液相冷媒を蒸発器１８に流出させる。
【００２８】
第２可変絞り弁１７は、圧縮機１１の低圧側になるように、気液分離器１６の下流側に設けられ、ＥＣＵ２２の制御信号に基づいて弁の開度が可変制御され、気液分離器１６からの液相冷媒を減圧させるとともに、蒸発器１８、エジェクタ１５への冷媒流量を調整し、ひいては、圧縮機１１への吸入圧力を調整する。
【００２９】
蒸発器１８は、図示しないファンの送風を受けて、冷媒を空気との熱交換により蒸発させる。
【００３０】
次に、本実施形態の作動を図２に示すモリエル線図に基づいて説明する。尚、このモリエル線図は、ヒートポンプ１０の動作点を描いたもので、図１のヒートポンプ１０上のａ〜ｋの状態が、図２に示すモリエル線図上のａ〜ｋに対応するものである。
【００３１】
圧縮機１１によって冷媒は、断熱的に圧縮され、等エントロピ線に沿って状態変化する。冷媒が給湯用水に対する充分な熱源となるように、冷媒の臨界点Ｋを超えるように圧縮し、高温となるようにしている。ここでは、エジェクタ１５により圧縮機１１への吸入圧力は従来の状態点ｊから状態点ｇに上昇されており、吐出圧力を従来と同一のＰ０とすると、吐出冷媒の温度は等温線図上でＴ１から温度の低いＴ２になるため、給湯用水との熱交換能力を確保するために、冷媒温度がＴ１となるように吐出圧力をＰ１に設定している。
【００３２】
オイル分離器１２でオイルと分離され、高温、高圧に圧縮された冷媒（状態点ａ）は、水熱交換器１３に流入し、給湯用水と熱交換し、給湯用水を昇温させ、冷媒自身は冷却される。（状態点ｂ）
この冷却された冷媒は、第１可変絞り弁１４を経て、エジェクタ１５のノズル１５ａで絞りロスを伴わないように等エントロピ線に沿って減圧、加速される。（状態点ｃ）加速された冷媒の流速によって蒸発器１８からの冷媒がエジェクタ１５の吸引部１５ｂから吸引され、（状態点ｄ）、水熱交換器１３、蒸発器１８の両者からの冷媒が混合部１５ｃで混合し、冷媒圧力はＰｍｉｘに上昇し（状態点ｅ）、更に、ディフューザ１５ｄで冷媒は減速され圧力Ｐｄ（中間圧力）まで上昇する。（状態点ｆ）
エジェクタ１５から流出した冷媒は、気液分離器１６に流入し、気液分離された気相冷媒は圧縮機１１に吸入される。（状態点ｇ）また、液相冷媒（状態点ｈ）は第２可変絞り弁１７を経て減圧され（状態点ｉ）、蒸発器１８に流入し外部空気との熱交換により蒸発され（状態点ｊ）、上記エジェクタ１５の吸引部１５ｂに吸引される。（状態点ｋ）
次に、本ヒートポンプ１０における高圧側圧力の制御について図３〜図５を用いて説明する。
【００３３】
給湯用水を所定の温度に昇温させるために熱交換能力を確保し、常に給湯用水を一定の温度に保つために効率的な熱交換を行なうには、水熱交換器１３に流入する給湯用水温度Ｔｗと、水熱交換器１３から流出する冷媒温度Ｔｒとの温度差ΔＴｘを所定の値に常に保つことが必要であり、水用温度センサ２０、冷媒用温度センサ２１によって検出される温度信号から、ＥＣＵ２２が第１可変絞り弁１４の弁開度を可変させることで、圧縮機１１の高圧側の圧力を制御している。因みに、温度差ΔＴｘが大きいと給湯用水の昇温が不足であり、温度差ΔＴｘが小さいと給湯用水の昇温が過度であり熱エネルギーを余分に使用していることになる。
【００３４】
熱交換能力、またその効率が最適となる給湯用水温度Ｔｗと冷媒温度Ｔｒとの温度差を目標温度差ΔＴとする。（本実施形態では８〜１０℃としている）今、ヒートポンプ１０内の諸条件の変動で、温度差ΔＴｘが目標温度差ΔＴよりも大きいとすると、ＥＣＵ２２は、第１可変絞り弁１４の弁開度を小さくなるように可変させ、高圧側の圧力を上げるように制御する。
【００３５】
図３は冷媒と給湯用水のエンタルピに対する温度の関係を等圧線図上で示したものである。実線が制御前、破線が制御後を示しており、給湯用水温度Ｔｗと冷媒温度Ｔｒ１との温度差ΔＴｘが目標温度差ΔＴに対して大きい状態にある時に、高圧側の圧力を上昇させることで、冷媒、給湯用水とも温度特性が破線に示すように上にずれ、この時、給湯用水温度Ｔｗが同一線上において冷媒の温度はＴｒ２に低下し、目標温度差ΔＴとなるようにしている。
【００３６】
逆に、温度差ΔＴｘが目標温度差ΔＴよりも小さいとすると、ＥＣＵ２２は、第１可変絞り弁１４の弁開度を大きくなるように可変させ、高圧側の圧力を下げるように制御する。
【００３７】
図４に示すように、給湯用水温度Ｔｗと冷媒温度Ｔｒ３との温度差ΔＴｘが目標温度差ΔＴに対して小さい状態にある時に、高圧側の圧力を低下させることで、冷媒、給湯用水とも温度特性が破線に示すように下にずれ、この時、給湯用水温度Ｔｗが同一線上において冷媒の温度はＴｒ２に上昇し、目標温度差ΔＴとなるようにしている。
【００３８】
更に、年間を通して夏場、冬場では当然外気温度に大きく差が生じ、圧縮機１１に吸入される冷媒の温度も変化する。吸入冷媒の温度が変化すれば、それに伴い吸入圧力も変化し、ひいては吐出圧力が変化する。例えば、夏場では冷媒温度が高くなり吐出圧力も高くなるが、この時ヒートポンプ１０の流路内の耐圧上の許容圧力を超えるような虞も考えられ、また、冬場では冷媒温度が低くなり吸入圧力も低くなるが、その分、圧縮機１１の圧縮仕事が増加する。
【００３９】
これを防止するために、ＥＣＵ２２は、圧力センサ１９で検出される冷媒の吐出圧力に基づいて、第２可変絞り弁１７の開度を可変して高圧側の圧力を制御するようにしている。第２可変絞り弁１７の開度を小さくすると蒸発器１８からエジェクタ１５に流入する冷媒流量が減少し圧縮機１１への吸入圧力が低下する。また、第２可変絞り弁１７の開度を大きくすると蒸発器１８からエジェクタ１５に流入する冷媒流量が増加し圧縮機１１への吸入圧力が上昇する。
【００４０】
図５に示すように、中間季節における吸入圧力ｎ点に対して、夏場はｍ点から圧縮されることになり、熱交換能力を確保するために冷媒温度がＴｒｉになるように圧縮すると、当然中間季節における吐出圧力Ｐ１よりも高いＰ２となる。この吐出圧力を下げるために、第２可変絞り弁１７の開度を小さくして吸入圧力をｎ点に下げてヒートポンプ１０を作動させるようにしている。
【００４１】
また、逆に、中間季節における吸入圧力ｎ点に対して、冬場はｏ点から圧縮されることになり、熱交換能力を確保するために冷媒温度がＴｒｉになるように圧縮すると、圧縮仕事は、等エントロピ線に沿って成されるため、吸入圧力が低い点ほど等エントロピ線図の勾配が小さくなるために、圧縮仕事Ｌ１に対して圧縮仕事Ｌ３に増加してしまう。この圧縮仕事Ｌ３を低減するために、第２可変絞り弁１７の開度を大きくして吸入圧力をｎ点に上げてヒートポンプ１０を作動させるようにしている。
【００４２】
以上、本発明のヒートポンプ１０の構成および作動より、以下の効果を得ることができる。
【００４３】
エジェクタ１５を設けて水熱交換器１３と蒸発器１８からの流出冷媒を混合して、圧縮機１１への吸入圧力を中間圧力になるように高めているので、図２に示すように圧縮仕事をＬ０からＬ１に低減できる。加えて、第1可変絞り弁１４で圧縮機１１の吐出圧力を可変し、吐出冷媒の温度を制御できるので、給湯用水との熱交換能力を常に確保できる。
【００４４】
また、目標温度差ΔＴが常に一定になるように第１可変絞り弁１４で高圧側圧力を制御するので、効率よく給湯用水を一定温度に昇温できる。
【００４５】
更に、外気温度の変化に対して第２可変絞り弁１７で吸入圧力を可変させ吐出圧力を制御するので、吸入圧力が高い場合には、吸入圧力を下げることにより吐出圧力を下げて、ヒートポンプサイクル内の耐圧上の許容圧力を超えない様に保護できる。また、吸入圧力が低い場合には吸入圧力を上げて、図５に示すように圧縮仕事をＬ３からＬ１に低減できる。
【００４６】
（その他の実施形態）
圧縮機１１への吸入圧力を制御するために、上記第１実施形態では冷媒の吐出圧力を検出する圧力センサ１９を用いたが、この吐出圧力に相関する物理量を用いて制御するようにしてもよい。例えば、水熱交換器１３の冷媒流出温度Ｔｒと外気温Ｔａを検出して吐出圧力に換算するようにしてもよい。
【００４７】
また、ヒートポンプ１０を給湯用水の熱源とする給湯器として説明したが、これに限らず、例えば、水熱交換器１３の負荷流体を空気に代えて、温風とする空調装置に用いてもよい。
【００４８】
更に、ヒートポンプ１０の冷媒は、ＣＯ2に限定されるものではなく、他のエチレン、エタン、酸化窒素等を用いてもよい。
【図面の簡単な説明】
【図１】本発明におけるヒートポンプの構成を示す模式図である。
【図２】ヒートポンプの作動を示すモリエル線図である。
【図３】温度差ΔＴｘを小さくする場合のＴ−ｈ線図（温度ーエンタルピ線図）上での説明図である。
【図４】温度差ΔＴｘを大きくする場合のＴ−ｈ線図（温度ーエンタルピ線図）上での説明図である。
【図５】外気温の差による吸入圧力を制御する場合のＰ−ｈ線図（圧力ーエンタルピ線図）上での説明図である。
【図６】従来技術のヒートポンプの構成を示す模式図である。
【符号の説明】
１０ ヒートポンプ
１１ 圧縮機
１３ 水熱交換器（高圧側熱交換器）
１４ 第１可変絞り弁
１５ エジェクタ
１６ 気液分離器
１７ 第２可変絞り弁
１８ 蒸発器
２２ ＥＣＵ（制御手段）[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a heat pump operating in a supercritical heat pump cycle, which is suitable for use as a heating means for heating hot water for a hot water heater, for example.
[0002]
[Prior art]
2. Description of the Related Art Conventionally, for example, a heat pump using a supercritical heat pump cycle that pressurizes a refrigerant pressure on a high pressure side to be equal to or higher than a critical pressure to heat hot water for a hot water heater is known.
[0003]
As shown in FIG. 6, the heat pump 10 includes a compressor 11, a water heat exchanger 13, an expansion valve 23, an evaporator 18, and a gas-liquid separator 16 that are sequentially connected by refrigerant piping. Hot water is supplied as a load fluid, and the hot water is heated by exchanging heat with a high-temperature and high-pressure refrigerant. The hot water supply water heated by the water heat exchanger 13 is stored in the tank 2 and is supplied to the user after the temperature is adjusted during use.
[0004]
The heat pump 10 is configured to control the high pressure of the cycle according to a temperature difference between hot water flowing into the water heat exchanger 13 and refrigerant flowing out of the water heat exchanger 13 in order to operate the heat pump cycle in a region where the cycle efficiency is high. Has been done.
[0005]
[Problems to be solved by the invention]
By the way, in order to further improve the cycle efficiency of the heat pump 10, the present inventors have devised that the suction pressure to the compressor 11 is increased using an ejector to reduce the compression work of the compressor 11. In this case, since the refrigerant temperature on the discharge side of the compressor 11 is lowered, there is a problem that the heat exchange capability in the water heat exchanger 13 is lowered.
[0006]
In view of the above problems, an object of the present invention is to provide a heat pump capable of reducing the compression work of a compressor in a heat pump cycle and ensuring the heat exchange capability of a water heat exchanger.
[0007]
[Means for Solving the Problems]
In order to achieve the above object, the present invention employs the following technical means.
[0008]
In the first aspect of the invention, the gas-phase refrigerant sucked from the gas-liquid separator (16) is pressurized to a critical pressure or higher by the compressor (11), and the refrigerant and the load fluid are pressurized by the high-pressure side heat exchanger (13). In the heat pump that exchanges heat as a counter flow, the refrigerant that flows out from the high-pressure side heat exchanger (13) and the refrigerant that flows out from the evaporator (18) that evaporates the liquid-phase refrigerant in the gas-liquid separator (16) After sucking and depressurizing to an intermediate pressure of the compressor (11), an ejector (15) that flows into the gas-liquid separator (16) and a high- pressure side heat exchanger (on the high-pressure side of the compressor (11)) provided downstream of 13), provided with the refrigerant flow first variable throttle valve for variably (14), the low pressure side of the compressor (the so as to adjust the suction pressure to 11) compressor (11) is, the second variable throttle valve for varying the coolant flow rate and (17) Only, the high pressure side of the first, second variable throttle valve (14, 17) opening the control means for controlling (22) of the compressor to ensure the heat exchange capacity of the load fluid (11) The second variable throttle so as to control the pressure on the high pressure side of the compressor (11) by controlling the opening of the first variable throttle valve (14) in order to change the pressure of the compressor and adjusting the suction pressure. It is characterized in that the opening degree of the valve (17) is controlled.
[0009]
Thereby, since the suction pressure of the compressor (11) is raised by the ejector (15), the compression work can be reduced. In addition, since the discharge pressure of the compressor (11) can be varied by the first variable throttle valve (14) and the temperature of the discharged refrigerant can be controlled, the heat exchange capability with the load fluid can always be ensured.
[0010]
In the second aspect of the invention, the temperature difference between the load fluid flowing into the high-pressure side heat exchanger (13) and the refrigerant flowing out from the high-pressure side heat exchanger (13) is controlled by the control means (22). The opening degree of the first variable throttle valve (14) is controlled so as to be ΔT.
[0011]
Thus, the target temperature difference ΔT is controlled so as to be always constant, so that the load fluid can be efficiently heated to a constant temperature.
[0012]
In the invention according to claim 3, the opening of the second variable throttle valve (17) is controlled by the control means (22) in accordance with the discharge pressure of the compressor (11) or a physical quantity correlated with the discharge pressure. It is characterized by doing.
[0013]
Thereby, for example, since the suction pressure is controlled by changing the suction pressure with the second variable throttle valve in response to a change in the outside air temperature, if the suction pressure is high, the discharge pressure is lowered by lowering the suction pressure, It can be protected so that it does not exceed the allowable pressure in the heat pump cycle. Further, when the suction pressure is low, the compression work can be reduced by increasing the suction pressure.
[0014]
The invention according to claim 4 is characterized in that the second variable throttle valve (17) adjusts the flow rate of the refrigerant flowing into the ejector (15) from the evaporator (18). Further, the invention according to claim 5 is characterized in that the second variable throttle valve (17) is provided on the downstream side of the gas-liquid separator (16). In the invention according to claim 6, the control means (22) controls the opening of the first variable throttle valve (14) so that the refrigerant temperature becomes Tri in order to ensure heat exchange capability. When the discharge pressure from the compressor (11) reaches P2, which is higher than P1, the opening of the second variable throttle valve (17) is reduced, and the refrigerant temperature becomes Tri to ensure heat exchange capability. Thus, when the discharge pressure from the compressor (11) becomes P3 lower than P1 by controlling the opening of the first variable throttle valve (14), the opening of the second variable throttle valve (17) is increased. It is characterized by doing. In addition, the code | symbol in the bracket | parenthesis of each said means shows a corresponding relationship with the specific means of embodiment description mentioned later.
[0015]
DETAILED DESCRIPTION OF THE INVENTION
(First embodiment)
A first embodiment of the present invention is shown in FIG. 1st Embodiment is applied to the heat pump type water heater 1 which uses the heat pump 10 of this invention as a heat source.
[0016]
The heat pump type hot water heater 1 is a system in which heated hot water supply water is stored in a tank 2, hot water supply water is taken out from the tank 2 during use, and the temperature is adjusted and supplied to a user.
[0017]
The tank 2 is made of metal (for example, made of stainless steel) excellent in corrosion resistance and has a heat insulating structure, and can keep hot hot water for a long time. In addition, although the hot water supply water stored in the tank 2 may be used directly in a kitchen, a bath, or the like, it can be used as a heat source other than for hot water supply, for example, for floor heating or indoor air conditioning.
[0018]
The electric pump 3 is provided in a hot water pipe 4 that annularly connects the tank 2 and a high-pressure side heat exchanger (hereinafter referred to as a water heat exchanger) 13 of the heat pump 10, and the tank 2, the water heat exchanger 13, The hot water supply water is circulated between the two, and the amount of circulating water is adjusted according to the rotation speed of the built-in motor.
[0019]
The heat pump 10 is configured such that the refrigerant pressure on the high-pressure side becomes equal to or higher than the critical pressure in order to sufficiently secure the refrigerant temperature (130 ° C. in the present embodiment) as a heat source for the hot water supply water, and the compressor 11, the water heat exchanger 13, an ejector 15, a gas-liquid separator 16, an evaporator 18, and the like, and carbon dioxide (CO 2 ) having a low critical pressure is used as a refrigerant.
[0020]
The compressor 11 is driven by a built-in electric motor (not shown), and compresses and discharges the gas-phase refrigerant sucked from a gas-liquid separator 16 described later to a critical pressure or more.
[0021]
The oil separator 12 is provided between the compressor 11 and the water heat exchanger 13, separates the refrigerant and oil discharged from the compressor 11, stores excess oil, and sequentially returns the oil to the compressor 11. Only the separated refrigerant flows into the water heat exchanger 13.
[0022]
The water heat exchanger 13 exchanges heat between the high-pressure gas-phase refrigerant discharged from the compressor 11 and hot water supply water as a load fluid so that the flow direction of the refrigerant and the flow direction of the hot water supply face each other. It is configured.
[0023]
And the pressure sensor 19 which detects the discharge pressure of the refrigerant | coolant discharged from the compressor 11 in the refrigerant | coolant inflow side and the outflow side of the water heat exchanger 13, respectively, and the temperature of the hot water supply water which flows in into the water heat exchanger 13 are detected. The temperature sensor for water 20 and the temperature sensor for refrigerant 21 for detecting the temperature of the refrigerant flowing out from the water heat exchanger 13 are provided, and the detection signals of the sensors 19 to 21 are electronic control units (hereinafter referred to as ECU) 22. Is input.
[0024]
ECU22 controls the opening degree of the 1st, 2nd variable throttle valves 14 and 17 mentioned later so that heat pump 10 can be operated efficiently.
[0025]
The first variable throttle valve 14 is provided on the downstream side of the water heat exchanger 13 so as to be on the high pressure side of the compressor 11, and the opening degree of the valve is variably controlled based on the control signal of the ECU 22. Adjust the discharge pressure.
[0026]
Furthermore, an ejector 15 including a nozzle 15a, a suction unit 15b, a mixing unit 15c, and a diffuser 15d is provided on the downstream side of the first variable throttle valve 14. The ejector 15 reduces the refrigerant heat-exchanged by the water heat exchanger 13 to the intermediate pressure of the compressor 11. First, the refrigerant flowing out of the water heat exchanger 13 is accelerated under reduced pressure without any squeezing loss by the nozzle 15a that contracts smoothly. Next, due to the suction effect of the refrigerant flow accelerated by the nozzle 15a, the refrigerant flowing out from the evaporator 18, which will be described later, is sucked from the suction portion 15b, mixed in the mixing portion 15c, and the pressure is increased, so that the diffuser expands smoothly. Decelerated by 15d and increased to intermediate pressure.
[0027]
The gas-liquid separator 16 gas-liquid separates the refrigerant flowing out from the ejector 15 to store excess refrigerant in the cycle, sucks the gas-phase refrigerant into the compressor 11, and flows out the liquid-phase refrigerant into the evaporator 18. Let
[0028]
The second variable throttle valve 17 is provided on the downstream side of the gas-liquid separator 16 so as to be on the low pressure side of the compressor 11, and the opening degree of the valve is variably controlled based on the control signal of the ECU 22, so that the gas-liquid separation is performed. The pressure of the liquid-phase refrigerant from the compressor 16 is reduced, the refrigerant flow rate to the evaporator 18 and the ejector 15 is adjusted, and consequently the suction pressure to the compressor 11 is adjusted.
[0029]
The evaporator 18 receives air from a fan (not shown) and evaporates the refrigerant by heat exchange with air.
[0030]
Next, the operation of the present embodiment will be described based on the Mollier diagram shown in FIG. This Mollier diagram depicts the operating point of the heat pump 10, and the states a to k on the heat pump 10 in FIG. 1 correspond to a to k on the Mollier diagram shown in FIG. is there.
[0031]
The refrigerant is adiabatically compressed by the compressor 11 and changes its state along the isentropic line. The refrigerant is compressed so as to exceed the critical point K of the refrigerant so that the refrigerant becomes a sufficient heat source for the hot water supply water so that the temperature becomes high. Here, the suction pressure to the compressor 11 by the ejector 15 is increased from the conventional state point j to the state point g, and when the discharge pressure is P0 which is the same as the conventional one, the temperature of the discharged refrigerant is on the isotherm. Since the temperature is changed from T1 to T2, which is lower in temperature, the discharge pressure is set to P1 so that the refrigerant temperature becomes T1 in order to ensure heat exchange capability with hot water.
[0032]
The refrigerant (state point a) separated from oil by the oil separator 12 and compressed to a high temperature and high pressure flows into the water heat exchanger 13 to exchange heat with the hot water supply water, raise the temperature of the hot water supply water, and the refrigerant itself. Is cooled. (State point b)
This cooled refrigerant passes through the first variable throttle valve 14 and is depressurized and accelerated along the isentropic line so as not to cause a throttle loss by the nozzle 15a of the ejector 15. (State point c) The refrigerant from the evaporator 18 is sucked from the suction part 15b of the ejector 15 by the flow rate of the accelerated refrigerant, and (state point d), the refrigerant from both the water heat exchanger 13 and the evaporator 18 is sucked. The refrigerant is mixed in the mixing unit 15c, the refrigerant pressure rises to Pmix (state point e), and further, the refrigerant is decelerated by the diffuser 15d and rises to the pressure Pd (intermediate pressure). (State point f)
The refrigerant that has flowed out of the ejector 15 flows into the gas-liquid separator 16, and the gas-phase separated refrigerant is sucked into the compressor 11. (State point g) The liquid refrigerant (state point h) is depressurized via the second variable throttle valve 17 (state point i), flows into the evaporator 18 and is evaporated by heat exchange with external air (state point). j) Suction is performed by the suction portion 15b of the ejector 15. (State point k)
Next, control of the high-pressure side pressure in the heat pump 10 will be described with reference to FIGS.
[0033]
In order to ensure heat exchange capacity to raise the temperature of the hot water supply water to a predetermined temperature and to perform efficient heat exchange in order to always maintain the hot water supply water at a constant temperature, the hot water supply water flowing into the water heat exchanger 13 It is necessary to always keep the temperature difference ΔTx between the temperature Tw and the refrigerant temperature Tr flowing out of the water heat exchanger 13 at a predetermined value, and the temperature signal detected by the water temperature sensor 20 and the refrigerant temperature sensor 21 is a temperature signal. Therefore, the ECU 22 controls the pressure on the high pressure side of the compressor 11 by changing the valve opening degree of the first variable throttle valve 14. Incidentally, if the temperature difference ΔTx is large, the temperature rise of the hot water supply water is insufficient, and if the temperature difference ΔTx is small, the temperature rise of the hot water supply water is excessive and excessive heat energy is used.
[0034]
A temperature difference between the hot water supply water temperature Tw and the refrigerant temperature Tr at which the heat exchange capacity and efficiency are optimized is defined as a target temperature difference ΔT. (In this embodiment, the temperature is set to 8 to 10 ° C.) Now, assuming that the temperature difference ΔTx is larger than the target temperature difference ΔT due to changes in various conditions in the heat pump 10, the ECU 22 opens the first variable throttle valve 14. The pressure is varied so as to decrease, and control is performed to increase the pressure on the high pressure side.
[0035]
FIG. 3 shows the relationship of temperature to the enthalpy of refrigerant and hot water supply water on an isobaric diagram. The solid line indicates before control, and the broken line indicates after control. When the temperature difference ΔTx between the hot water supply water temperature Tw and the refrigerant temperature Tr1 is larger than the target temperature difference ΔT, the pressure on the high pressure side is increased. The temperature characteristics of both the refrigerant and the hot water supply water are shifted upward as shown by the broken line. At this time, the temperature of the refrigerant falls to Tr2 when the hot water supply water temperature Tw is on the same line, so that the target temperature difference ΔT is obtained.
[0036]
On the other hand, if the temperature difference ΔTx is smaller than the target temperature difference ΔT, the ECU 22 controls the first variable throttle valve 14 so that the valve opening degree is increased and the pressure on the high pressure side is decreased.
[0037]
As shown in FIG. 4, when the temperature difference ΔTx between the hot water supply water temperature Tw and the refrigerant temperature Tr3 is smaller than the target temperature difference ΔT, the pressure on the high-pressure side is decreased, so that the temperature of both the refrigerant and the hot water supply water is reduced. The characteristic shifts downward as indicated by the broken line, and at this time, the temperature of the hot water supply water Tw is on the same line, the temperature of the refrigerant rises to Tr2, and becomes the target temperature difference ΔT.
[0038]
Further, naturally, a large difference occurs in the outside air temperature in summer and winter throughout the year, and the temperature of the refrigerant sucked into the compressor 11 also changes. If the temperature of the suction refrigerant changes, the suction pressure changes accordingly, and the discharge pressure changes accordingly. For example, in summer, the refrigerant temperature becomes high and the discharge pressure becomes high. At this time, there is a possibility that the allowable pressure in the flow path of the heat pump 10 may be exceeded. In winter, the refrigerant temperature becomes low and the suction pressure becomes low. However, the compression work of the compressor 11 increases accordingly.
[0039]
In order to prevent this, the ECU 22 controls the pressure on the high pressure side by varying the opening of the second variable throttle valve 17 based on the refrigerant discharge pressure detected by the pressure sensor 19. If the opening degree of the second variable throttle valve 17 is reduced, the flow rate of the refrigerant flowing from the evaporator 18 into the ejector 15 is reduced, and the suction pressure to the compressor 11 is reduced. Further, when the opening of the second variable throttle valve 17 is increased, the flow rate of the refrigerant flowing from the evaporator 18 into the ejector 15 is increased, and the suction pressure to the compressor 11 is increased.
[0040]
As shown in FIG. 5, the summer pressure is compressed from the point m with respect to the suction pressure n point in the intermediate season, and naturally when the refrigerant temperature is compressed to Tri in order to ensure the heat exchange capability. P2 is higher than the discharge pressure P1 in the intermediate season. In order to reduce the discharge pressure, the opening of the second variable throttle valve 17 is reduced to reduce the suction pressure to the point n, thereby operating the heat pump 10.
[0041]
On the other hand, with respect to the suction pressure n point in the middle season, in the winter season, the compression is performed from the point o. If the refrigerant temperature is compressed to Tri to ensure heat exchange capacity, the compression work is Since it is formed along the isentropic curve, the lower the suction pressure, the smaller the gradient of the isentropic diagram. In order to reduce this compression work L3, the opening degree of the second variable throttle valve 17 is increased to increase the suction pressure to the n point, and the heat pump 10 is operated.
[0042]
As described above, the following effects can be obtained from the configuration and operation of the heat pump 10 of the present invention.
[0043]
Since the ejector 15 is provided and the refrigerant flowing out from the water heat exchanger 13 and the evaporator 18 is mixed to increase the suction pressure to the compressor 11 so that it becomes an intermediate pressure, as shown in FIG. Can be reduced from L0 to L1. In addition, since the discharge pressure of the compressor 11 can be varied by the first variable throttle valve 14 and the temperature of the discharged refrigerant can be controlled, the heat exchange capability with hot water supply water can always be ensured.
[0044]
Moreover, since the high pressure side pressure is controlled by the first variable throttle valve 14 so that the target temperature difference ΔT is always constant, the hot water supply water can be efficiently heated to a constant temperature.
[0045]
Further, since the suction pressure is varied by the second variable throttle valve 17 to control the discharge pressure in response to a change in the outside air temperature, when the suction pressure is high, the discharge pressure is lowered by lowering the suction pressure, and the heat pump cycle. It can be protected so as not to exceed the allowable pressure on the inside pressure resistance. Further, when the suction pressure is low, the suction pressure can be increased to reduce the compression work from L3 to L1, as shown in FIG.
[0046]
(Other embodiments)
In order to control the suction pressure to the compressor 11, the pressure sensor 19 that detects the discharge pressure of the refrigerant is used in the first embodiment. However, the control may be performed using a physical quantity that correlates with the discharge pressure. Good. For example, the refrigerant outflow temperature Tr and the outside air temperature Ta of the water heat exchanger 13 may be detected and converted into discharge pressure.
[0047]
Moreover, although demonstrated as a water heater which uses the heat pump 10 as a heat source of hot water for water supply, it is not restricted to this, For example, it may replace with the load fluid of the water heat exchanger 13, and may be used for the air-conditioner which uses warm air. .
[0048]
Furthermore, the refrigerant of the heat pump 10 is not limited to CO 2 , and other ethylene, ethane, nitrogen oxide, or the like may be used.
[Brief description of the drawings]
FIG. 1 is a schematic diagram showing a configuration of a heat pump in the present invention.
FIG. 2 is a Mollier diagram showing the operation of the heat pump.
FIG. 3 is an explanatory diagram on a Th diagram (temperature-enthalpy diagram) when the temperature difference ΔTx is reduced.
FIG. 4 is an explanatory diagram on a Th diagram (temperature-enthalpy diagram) when the temperature difference ΔTx is increased.
FIG. 5 is an explanatory diagram on a Ph diagram (pressure-enthalpy diagram) in the case of controlling the suction pressure due to a difference in outside air temperature.
FIG. 6 is a schematic diagram showing a configuration of a conventional heat pump.
[Explanation of symbols]
10 Heat pump 11 Compressor 13 Water heat exchanger (High pressure side heat exchanger)
14 First variable throttle valve 15 Ejector 16 Gas-liquid separator 17 Second variable throttle valve 18 Evaporator 22 ECU (control means)

Claims (6)

サイクル内の冷媒を気液分離する気液分離器（１６）と、
この気液分離器（１６）より吸入した気相冷媒を臨界圧力以上に加圧して吐出する圧縮機（１１）と、
この圧縮機（１１）で加圧された冷媒と負荷流体とを対向流として熱交換させる高圧側熱交換器（１３）とを有するヒートポンプにおいて、
前記高圧側熱交換器（１３）から流出する冷媒、および、前記気液分離器（１６）の液相冷媒を蒸発させる蒸発器（１８）から流出する冷媒を吸入し、前記圧縮機（１１）の中間圧力まで減圧した後、前記気液分離器（１６）に流入させるエジェクタ（１５）と、
前記圧縮機（１１）の高圧側となるように前記高圧側熱交換器（１３）の下流側に設けられ、冷媒流量を可変する第１の可変絞り弁（１４）と、
前記圧縮機（１１）への吸入圧力を調整するように前記圧縮機（１１）の低圧側に設けられ、冷媒流量を可変する第２の可変絞り弁（１７）と、
前記第１、第２の可変絞り弁（１４、１７）の開度を制御する制御手段（２２）とを設け、
前記制御手段（２２）は、前記負荷流体との熱交換能力を確保するように前記圧縮機（１１）の高圧側の圧力を変化させるために前記第１の可変絞り弁（１４）の開度を制御し、前記吸入圧力を調整することで前記圧縮機（１１）の高圧側の圧力を制御するように前記第２の可変絞り弁（１７）の開度を制御するようにしたことを特徴とするヒートポンプ。A gas-liquid separator (16) for gas-liquid separation of the refrigerant in the cycle;
A compressor (11) that pressurizes and discharges the gas-phase refrigerant sucked from the gas-liquid separator (16) to a critical pressure or higher;
In a heat pump having a high-pressure side heat exchanger (13) for exchanging heat between the refrigerant pressurized by the compressor (11) and the load fluid as a counter flow,
The refrigerant that flows out from the high-pressure side heat exchanger (13) and the refrigerant that flows out from the evaporator (18) that evaporates the liquid-phase refrigerant in the gas-liquid separator (16) are sucked into the compressor (11). And an ejector (15) that flows into the gas-liquid separator (16).
A first variable throttle valve (14) provided on the downstream side of the high-pressure side heat exchanger (13) so as to be on the high-pressure side of the compressor (11) , wherein the refrigerant flow rate is variable;
A second variable throttle valve (17) that is provided on the low pressure side of the compressor (11) so as to adjust the suction pressure to the compressor (11) and varies the refrigerant flow rate;
Control means (22) for controlling the opening degree of the first and second variable throttle valves (14, 17);
The control means (22) opens the opening of the first variable throttle valve (14) to change the pressure on the high pressure side of the compressor (11) so as to ensure heat exchange capability with the load fluid. And the opening of the second variable throttle valve (17) is controlled so as to control the pressure on the high pressure side of the compressor (11) by adjusting the suction pressure. And heat pump. 前記制御手段（２２）は、前記高圧側熱交換器（１３）に流入する負荷流体と前記高圧側熱交換器（１３）から流出する冷媒との温度差が目標温度差ΔＴとなるように、
前記第１の可変絞り弁（１４）の開度を制御することを特徴とする請求項1に記載のヒートポンプ。The control means (22) is configured so that the temperature difference between the load fluid flowing into the high-pressure side heat exchanger (13) and the refrigerant flowing out from the high-pressure side heat exchanger (13) becomes a target temperature difference ΔT.
The heat pump according to claim 1, wherein the opening degree of the first variable throttle valve (14) is controlled. 前記制御手段（２２）は、前記圧縮機（１１）の吐出圧力、あるいは吐出圧力に相関する物理量に応じて、前記第２の可変絞り弁（１７）の開度を制御することを特徴とする請求項１または請求項２に記載のヒートポンプ。  The control means (22) controls the opening of the second variable throttle valve (17) according to the discharge pressure of the compressor (11) or a physical quantity correlated with the discharge pressure. The heat pump according to claim 1 or 2. 前記第２の可変絞り弁（１７）は、前記蒸発器（１８）から前記エジェクタ（１５）に流入する冷媒流量を調整することを特徴とする請求項１〜請求項３のいずれか１つに記載のヒートポンプ。The said 2nd variable throttle valve (17) adjusts the refrigerant | coolant flow volume which flows in into the said ejector (15) from the said evaporator (18), Any one of Claims 1-3 characterized by the above-mentioned. The heat pump described. 前記第２の可変絞り弁（１７）は、前記気液分離器（１６）の下流側に設けられていることを特徴とする請求項１〜請求項４のいずれか１つに記載のヒートポンプ。The heat pump according to any one of claims 1 to 4, wherein the second variable throttle valve (17) is provided downstream of the gas-liquid separator (16). 前記制御手段（２２）は、前記熱交換能力を確保するために冷媒温度がＴｒｉになるように前記第１の可変絞り弁（１４）の開度を制御することで前記圧縮機（１１）からの吐出圧力がＰ１より高いＰ２になると、前記第２の可変絞り弁（１７）の開度を小さくし、The control means (22) controls the opening of the first variable throttle valve (14) so that the refrigerant temperature becomes Tri in order to ensure the heat exchanging capacity. When the discharge pressure of P2 is higher than P1, the opening of the second variable throttle valve (17) is reduced,
また、前記熱交換能力を確保するために冷媒温度がＴｒｉになるように前記第１の可変絞り弁（１４）の開度を制御することで前記圧縮機（１１）からの吐出圧力がＰ１より低いＰ３になると、前記第２の可変絞り弁（１７）の開度を大きくすることを特徴とする請求項１〜請求項５のいずれか１つに記載のヒートポンプ。  Further, by controlling the opening of the first variable throttle valve (14) so that the refrigerant temperature becomes Tri in order to ensure the heat exchange capability, the discharge pressure from the compressor (11) is less than P1. The heat pump according to any one of claims 1 to 5, wherein the opening of the second variable throttle valve (17) is increased when P3 is low.
JP2000264861A 2000-09-01 2000-09-01 heat pump Expired - Fee Related JP3835141B2 (en)

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JP2009133624A (en) * 2005-03-14 2009-06-18 Mitsubishi Electric Corp Refrigerating/air-conditioning device
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