JP2005249384A - Refrigerating cycle device - Google Patents

Refrigerating cycle device Download PDF

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JP2005249384A
JP2005249384A JP2005112540A JP2005112540A JP2005249384A JP 2005249384 A JP2005249384 A JP 2005249384A JP 2005112540 A JP2005112540 A JP 2005112540A JP 2005112540 A JP2005112540 A JP 2005112540A JP 2005249384 A JP2005249384 A JP 2005249384A
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heat exchanger
compressor
refrigeration cycle
high pressure
refrigerant
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Takashi Okazaki
多佳志 岡崎
Toshihiko Enomoto
寿彦 榎本
Yoshihiro Sumida
嘉裕 隅田
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Mitsubishi Electric Corp
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Mitsubishi Electric 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
    • F25B13/00—Compression machines, plants or systems, with reversible cycle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • 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
    • 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
    • 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
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/06—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point using expanders

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  • Air Conditioning Control Device (AREA)

Abstract

<P>PROBLEM TO BE SOLVED: To save energy by always maintaining a high coefficient of performance in a refrigerating cycle device with a refrigerant operated at a supercritical pressure on a high-pressure side. <P>SOLUTION: This refrigerating cycle device comprises a memory 10a holding a relation between the outlet pipe temperature of a high-pressure side heat exchanger (outdoor heat exchanger 2 in cooling, indoor heat exchanger 6 in heating) and a high-pressure range in which the coefficient of the performance of a refrigerating cycle is maximized, pipe temperature detectors 20 to 23 detecting the outlet pipe temperature of the high-pressure side heat exchanger or the temperatures of pipes starting at a compressor 1 to the high-pressure side heat exchanger, and a controller 10 controlling, in a high-pressure range, the opening of an electronic expansion valve, the rotational speed of the compressor 1, and the rotational speed of an indoor fan 7 or an outdoor fan 3 based on the pipe temperatures detected by the pipe temperature detectors 20 to 23 and the optimum relation in the memory. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

この発明は、高圧側が超臨界圧力で運転される冷媒を用いた空気調和機やヒートポンプなどの冷凍サイクル装置およびその制御方法に関し、特に超臨界圧力で運転される冷媒の特徴を有効に利用し、高効率な運転を行って省エネルギー化を図ることができる冷凍サイクル装置に関するものである。   The present invention relates to a refrigeration cycle apparatus such as an air conditioner or a heat pump using a refrigerant whose high pressure side is operated at a supercritical pressure, and a control method thereof, in particular, effectively utilizing the characteristics of the refrigerant operated at a supercritical pressure, The present invention relates to a refrigeration cycle apparatus capable of energy saving by performing highly efficient operation.

近年、地球環境保護や機器効率の向上の観点から、空気調和機の冷媒として高圧側が超臨界圧力で運転される二酸化炭素などの超臨界冷媒の適用検討が行われている。高圧側が超臨界圧力で運転される冷媒を用いる空気調和機としては、たとえば、特許文献２に記載された超臨界蒸気圧縮サイクルの運転方法および運転装置がある。   In recent years, from the viewpoint of protecting the global environment and improving the efficiency of equipment, the application of a supercritical refrigerant such as carbon dioxide, which is operated at a supercritical pressure on the high pressure side, has been studied as a refrigerant for air conditioners. As an air conditioner using a refrigerant whose high pressure side is operated at a supercritical pressure, for example, there is a supercritical vapor compression cycle operation method and operation device described in Patent Document 2.

この超臨界蒸気圧縮サイクルの運転方法では、超臨界冷凍サイクル装置の能力を制御する技術が記載され、圧縮機、高圧側熱交換器（放熱器）、絞り手段および蒸発器を備えた超臨界冷凍サイクル装置において、予め、放熱器に流入して冷媒と熱交換する流体の流入温度と、冷凍サイクルの成績係数が最大となる最適高圧圧力との特性を求めておき、冷凍サイクルの制御量としてのこの流体の流入温度（たとえば、室内温度）を測定し、この測定値が目標値に達せず、能力を必要とする場合には、上述した流体の流入温度と高圧圧力との特性をもとに、絞り手段の開度を調整し、成績係数が最大となるような高圧圧力で運転し、一方、能力を必要としない場合には、絞り手段の開度を調整して、高圧圧力を予め設定した最低高圧圧力まで低下させて能力を減少させ、この最低高圧圧力による運転が予め設定した一定時間継続した場合に、圧縮機を停止させ、手動による圧縮機のオン／オフ制御を行うようにしている。   In this supercritical vapor compression cycle operation method, a technique for controlling the capability of the supercritical refrigeration cycle apparatus is described, and a supercritical refrigeration equipped with a compressor, a high-pressure side heat exchanger (heat radiator), a throttle means, and an evaporator. In the cycle device, the characteristics of the inflow temperature of the fluid that flows into the radiator and exchanges heat with the refrigerant and the optimum high pressure that maximizes the coefficient of performance of the refrigeration cycle are obtained in advance, and the control amount of the refrigeration cycle If the fluid inflow temperature (for example, the room temperature) is measured and the measured value does not reach the target value and the capacity is required, the characteristics of the fluid inflow temperature and the high pressure described above are used. Adjust the opening of the throttle means and operate at a high pressure that maximizes the coefficient of performance. On the other hand, if you do not need capacity, adjust the opening of the throttle means and set the high pressure in advance. Reduced to the lowest high pressure Thereby reducing the capacity, when the operation with the lowest high pressure is continued for a certain preset time, the compressor is stopped, and to perform on / off control of the compressor by hand.

ここで、高圧圧力が超臨界圧力となる冷媒としては、たとえば二酸化炭素が使用される。また、圧縮機が回転数可変型の場合、室温制御系と高圧制御系とは独立させて運転を行う。すなわち、能力の調整は、圧縮機回転数の調整によって行い、一方、絞り手段の開度を調整して、成績係数が最大となるような高圧圧力で運転するようにしている。   Here, for example, carbon dioxide is used as the refrigerant whose high pressure becomes the supercritical pressure. When the compressor is a variable speed type, the room temperature control system and the high pressure control system are operated independently. That is, the capacity is adjusted by adjusting the number of rotations of the compressor, while the opening of the throttle means is adjusted to operate at a high pressure that maximizes the coefficient of performance.

図１５は、上述した従来の超臨界蒸気圧縮サイクルを用いた空気調和機の構成を示す図である。図１５において、この空気調和機は、室内熱交換器６、室内ファン７を含む室内ユニットＡと、室外熱交換器２、室外ファン３、圧縮機１およびアキュムレータ１２を含む室外ユニットＢとを有し、室内ユニットＡと室外ユニットＢとは、接続配管（液配管）５および接続配管（ガス配管）８によって接続されている。   FIG. 15 is a diagram showing a configuration of an air conditioner using the above-described conventional supercritical vapor compression cycle. In FIG. 15, this air conditioner has an indoor unit A including an indoor heat exchanger 6 and an indoor fan 7, and an outdoor unit B including an outdoor heat exchanger 2, an outdoor fan 3, a compressor 1 and an accumulator 12. The indoor unit A and the outdoor unit B are connected by a connection pipe (liquid pipe) 5 and a connection pipe (gas pipe) 8.

圧縮機１は、吸入した冷媒ガスを圧縮して室外熱交換器２または室内熱交換器６に送り、アキュムレータ１２は、圧縮機１の吸入側に接続され室外熱交換器２または室内熱交換器６から戻る液冷媒を貯留する。冷房運転時には、四方切替弁９によって圧縮機１と室外熱交換器２とが連通し、室内熱交換器６および室外熱交換器２は、それぞれ蒸発器および放熱器として機能する。暖房運転時には、四方切替弁９によって圧縮機１と室内熱交換器６とが連通し、室内熱交換器６および室外熱交換器２は、それぞれ放熱器および蒸発器として機能する。   The compressor 1 compresses the sucked refrigerant gas and sends it to the outdoor heat exchanger 2 or the indoor heat exchanger 6, and the accumulator 12 is connected to the suction side of the compressor 1 and is connected to the outdoor heat exchanger 2 or the indoor heat exchanger. The liquid refrigerant returning from 6 is stored. During the cooling operation, the compressor 1 and the outdoor heat exchanger 2 communicate with each other by the four-way switching valve 9, and the indoor heat exchanger 6 and the outdoor heat exchanger 2 function as an evaporator and a radiator, respectively. During the heating operation, the compressor 1 and the indoor heat exchanger 6 communicate with each other by the four-way switching valve 9, and the indoor heat exchanger 6 and the outdoor heat exchanger 2 function as a radiator and an evaporator, respectively.

室内熱交換器６および室外熱交換器２は、室内空気および室外空気の温度をそれぞれ検出する室内温度検出器１５および室外温度検出器１４をそれぞれ備えている。室外ユニットＢ内において、吐出圧力検出器１１は、圧縮機１と四方切替弁９との間の配管に設けられ、圧縮機１によって吐出される冷媒の圧力（高圧圧力）を検出する。コントローラ１１０は、室内温度検出器１５、室外温度検出器１４、および吐出圧力検出器１１によってそれぞれ検出された検出値を入力し、これらの検出値をもとに、電子式膨張弁４の開度を制御する。   The indoor heat exchanger 6 and the outdoor heat exchanger 2 include an indoor temperature detector 15 and an outdoor temperature detector 14 that detect the temperatures of the indoor air and the outdoor air, respectively. In the outdoor unit B, the discharge pressure detector 11 is provided in a pipe between the compressor 1 and the four-way switching valve 9 and detects the pressure (high pressure) of the refrigerant discharged by the compressor 1. The controller 110 inputs detection values detected by the indoor temperature detector 15, the outdoor temperature detector 14, and the discharge pressure detector 11, and based on these detection values, the opening degree of the electronic expansion valve 4. To control.

また、コントローラ１１０には、予め、高圧側の熱交換器（冷房時は室内熱交換器６、暖房時は室外熱交換器２）に流入する空気の温度すなわち放熱器流入空気温度と、成績係数ＣＯＰ（Coefficient of Performance）が最大となる高圧圧力との特性が記憶されている。コントローラ１１０は、制御量としての室温と目標室温との偏差を求め、成績係数ＣＯＰが最大となる高圧圧力を決定する。高圧圧力は、吐出圧力検出器１１で検出された高圧圧力と目標とする高圧圧力との偏差に応じて、コントローラ１１０が電子式膨張弁４の開度を調整することによって制御される。室温は、冷房運転が指示されている場合、放熱器流入空気温度を室外温度とみなされ、暖房運転が指示されている場合、放熱器流入空気温度を室内温度とみなされ、コントローラ１１０が室温目標値と室内温度とを比較することによって、冷凍サイクルの能力が必要とされているか否かが判断される。   In addition, the controller 110 previously stores the temperature of the air flowing into the high-pressure side heat exchanger (the indoor heat exchanger 6 during cooling and the outdoor heat exchanger 2 during heating), that is, the radiator inflow air temperature, and the coefficient of performance. The characteristic of the high pressure at which COP (Coefficient of Performance) is maximized is stored. The controller 110 obtains the deviation between the room temperature as the control amount and the target room temperature, and determines the high pressure at which the coefficient of performance COP is maximized. The high pressure is controlled by the controller 110 adjusting the opening of the electronic expansion valve 4 in accordance with the deviation between the high pressure detected by the discharge pressure detector 11 and the target high pressure. When the cooling operation is instructed, the room temperature is regarded as the outdoor temperature of the radiator inflow air. When the heating operation is instructed, the temperature of the radiator inflow air is regarded as the room temperature, and the controller 110 sets the room temperature target. By comparing the value with the room temperature, it is determined whether the capacity of the refrigeration cycle is required.

このようにして、従来の超臨界冷凍サイクルの温調制御系では、高圧側の熱交換器に流入する流入流体の温度毎に、成績係数ＣＯＰが最大となる高圧圧力が存在する特性を利用し、能力を必要とする場合は、成績係数ＣＯＰを優先して、成績係数ＣＯＰが最大となる高圧圧力で運転させる。一方、能力を必要としない場合は、成績係数ＣＯＰを無視し、高圧圧力を低下させることによって能力を低下させる。ただし、長期間の低成績係数での運転を避けるため、最低高圧圧力を設定し、この最低高圧圧力での運転が一定時間継続した場合、圧縮機１を停止し、手動によるオン／オフ制御によって温調制御を行う。これによって、冷凍サイクルを効率よく運転できる。   In this way, the temperature control system of the conventional supercritical refrigeration cycle uses the characteristic that there is a high-pressure pressure that maximizes the coefficient of performance COP for each temperature of the inflow fluid flowing into the high-pressure side heat exchanger. When the ability is required, the coefficient of performance COP is given priority and the operation is performed at a high pressure at which the coefficient of performance COP is maximized. On the other hand, when the ability is not required, the performance coefficient COP is ignored, and the ability is lowered by lowering the high pressure. However, in order to avoid operation with a low coefficient of performance for a long period of time, if the minimum high pressure is set and the operation at this minimum high pressure continues for a certain period of time, the compressor 1 is stopped and manual on / off control is performed. Perform temperature control. Thereby, the refrigeration cycle can be operated efficiently.

特開２００２−３０５４０９号公報JP 2002-305409 A 特開平１１−２１１２５１号公報JP 11-2111251 A

しかしながら、上述した従来の空気調和機では、室内ユニットＡが複数台設けられる場合、室内ユニットＡの台数変化や外気温度変化に伴って負荷変動が生じ、この負荷変動が生じた場合、室外ファン３あるいは室内ファン７の風量が変化することから、たとえば空気のような流体の流入温度と成績係数が最大となる最適高圧圧力との特性（関係）も変化することになる。この場合、上述した従来の空気調和機では、この流体の流量変化に対応した適切な制御を行うことができず、効率的な運転を行うことができないという問題点があった。   However, in the above-described conventional air conditioner, when a plurality of indoor units A are provided, a load change occurs with a change in the number of indoor units A or a change in outside air temperature. When this load change occurs, the outdoor fan 3 Alternatively, since the air volume of the indoor fan 7 changes, the characteristic (relationship) between the inflow temperature of a fluid such as air and the optimum high pressure at which the coefficient of performance becomes maximum also changes. In this case, the conventional air conditioner described above has a problem in that it cannot perform appropriate control corresponding to the change in the flow rate of the fluid and cannot perform efficient operation.

また、上述した従来の空気調和機では、高圧圧力を検知する手段として、圧力を直接測定する圧力検知器を用いているが、この圧力検知器を用いることによって冷媒回路の構造を複雑かつ大型化し、冷媒回路自体にかかるコストが高くなるという問題点があった。   Further, in the conventional air conditioner described above, a pressure detector that directly measures the pressure is used as a means for detecting the high pressure, but the structure of the refrigerant circuit is complicated and enlarged by using this pressure detector. There has been a problem that the cost of the refrigerant circuit itself becomes high.

さらに、上述した空気調和機では、放熱器出口温度をできる限り低温にすることによって冷凍効果が増大し、高い成績係数ＣＯＰを得ることができるが、放熱器出口部と蒸発器出口部とを熱交換させる従来の技術ではその効果に限界が存在し、放熱器出口温度を十分に低下させることができず、一層、高い成績係数ＣＯＰを得ることができないという問題点があった。なお、超臨界冷凍サイクルを用いた冷凍サイクル装置では、常に成績係数ＣＯＰの高い装置の出現が要望されている。   Furthermore, in the air conditioner described above, the refrigeration effect is increased by making the radiator outlet temperature as low as possible, and a high coefficient of performance COP can be obtained, but the radiator outlet and the evaporator outlet are heated. In the conventional technique to be replaced, there is a limit to the effect, and the radiator outlet temperature cannot be sufficiently lowered, and a higher coefficient of performance COP cannot be obtained. In addition, in the refrigeration cycle apparatus using the supercritical refrigeration cycle, the appearance of an apparatus having a high coefficient of performance COP is always demanded.

この発明は上記に鑑みてなされたもので、高圧側が超臨界圧力で運転される冷媒を用いた冷凍サイクル装置であって、常に高い成績係数を維持させて省エネルギー化を図ることができる冷凍サイクル装置を得ることを目的とする。   The present invention has been made in view of the above, and is a refrigeration cycle apparatus using a refrigerant whose high pressure side is operated at a supercritical pressure, and can maintain energy efficiency by always maintaining a high coefficient of performance. The purpose is to obtain.

上記目的を達成するため、この発明にかかる冷凍サイクル装置は、少なくとも圧縮機、高圧側熱交換器、絞り手段、および低圧側熱交換器を液配管およびガス配管を用いて接続した冷媒回路を形成するとともに、前記高圧側熱交換器および前記低圧側熱交換器に対して被加熱媒体および被冷却媒体を搬送する被加熱媒体搬送手段および被冷却媒体搬送手段を有し、高圧側が超臨界圧力で運転される冷媒を用いた冷凍サイクル装置において、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持した保持手段と、前記高圧側熱交換器の出口配管温度を検出する検出手段と、前記検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御する制御手段とを備えたことを特徴とする。   To achieve the above object, a refrigeration cycle apparatus according to the present invention forms a refrigerant circuit in which at least a compressor, a high-pressure side heat exchanger, a throttling means, and a low-pressure side heat exchanger are connected using a liquid pipe and a gas pipe. And a heated medium conveying means and a cooled medium conveying means for conveying a medium to be heated and a medium to be cooled to the high pressure side heat exchanger and the low pressure side heat exchanger, and the high pressure side is at a supercritical pressure. In the refrigeration cycle apparatus using the refrigerant to be operated, the holding means holding the relationship between the outlet pipe temperature of the high pressure side heat exchanger and the high pressure range where the coefficient of performance of the refrigeration cycle is near the maximum, and the high pressure side heat Based on the relationship between the detection means for detecting the outlet pipe temperature of the exchanger, the detection result of the detection means, and the holding means, the opening degree of the throttle means, the rotational speed of the compressor, Characterized in that a control means for the medium flow rate, and the controls at least one of the flow rate of the cooling medium is controlled in the high pressure range.

この発明によれば、保持手段が、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持し、制御手段が、前記高圧側熱交換器の出口配管温度を検出する検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにし、被加熱媒体あるいは被冷却媒体の流量が変化した場合であっても高圧圧力範囲で運転できるようにしている。   According to this invention, the holding means holds the relationship between the outlet pipe temperature of the high pressure side heat exchanger and the high pressure range where the coefficient of performance of the refrigeration cycle is near the maximum, and the control means is the high pressure side heat exchange. Based on the detection result of the detection means for detecting the outlet pipe temperature of the compressor and the relationship between the holding means, the opening degree of the throttle means, the rotation speed of the compressor, the flow rate of the heated medium, and the cooled medium At least one of the flow rates is controlled to be within the high pressure range, so that the operation can be performed within the high pressure range even when the flow rate of the heating medium or the cooling medium changes.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記絞り手段は、冷媒を膨張させる膨張機構部で形成され、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結したことを特徴とする。   In the refrigeration cycle apparatus according to the next invention, in the above invention, the throttling means is formed by an expansion mechanism section that expands a refrigerant, and the output shaft of the expansion mechanism section is connected to the compressor and the heated medium transport means. The cooling medium transfer means is connected to at least one drive shaft.

この発明によれば、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結し、膨張機構部に連結する圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の駆動力の一部として、膨張機構部の回転力を利用するようにしている。   According to this invention, the output shaft of the expansion mechanism unit is connected to at least one drive shaft of the compressor, the heated medium transport unit, and the cooled medium transport unit, and the compression is coupled to the expansion mechanism unit. The rotational force of the expansion mechanism is used as part of the driving force of the machine, heated medium conveying means, and cooled medium conveying means.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記圧縮機から前記高圧側熱交換器に至る配管の温度を検出する配管温度検出手段をさらに備え、前記制御手段は、前記配管温度検出手段が検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御することを特徴とする。   The refrigeration cycle apparatus according to the next invention further comprises pipe temperature detection means for detecting the temperature of the pipe from the compressor to the high pressure side heat exchanger in the above invention, wherein the control means is the pipe temperature detection. The high pressure is estimated based on the temperature detected by the means, and based on the estimated high pressure, the detection result of the detection means, and the relationship between the holding means, the opening degree of the throttle means, the compressor Controlling at least one of the rotational speed, the flow rate of the medium to be heated, and the flow rate of the medium to be cooled to control the high pressure range.

この発明によれば、制御手段が、前記配管温度検出手段が検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしている。   According to this invention, the control means estimates the high pressure based on the temperature detected by the pipe temperature detection means, and has the relationship between the estimated high pressure, the detection result of the detection means, and the holding means. And controlling at least one of the opening degree of the throttle means, the rotational speed of the compressor, the flow rate of the heated medium, and the flow rate of the cooled medium to control the high pressure range. Yes.

つぎの発明にかかる冷凍サイクル装置は、少なくとも圧縮機、高圧側熱交換器、膨張機構部、および低圧側熱交換器を液配管およびガス配管を用いて接続した冷媒回路を形成するとともに、前記高圧側熱交換器および前記低圧側熱交換器に対して被加熱媒体および被冷却媒体を搬送する被加熱媒体搬送手段および被冷却媒体搬送手段を有し、高圧側が超臨界圧力で運転される冷媒を用いた冷凍サイクル装置において、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結したことを特徴とする。   The refrigeration cycle apparatus according to the next invention forms a refrigerant circuit in which at least a compressor, a high-pressure side heat exchanger, an expansion mechanism, and a low-pressure side heat exchanger are connected using a liquid pipe and a gas pipe, and the high pressure A refrigerant having heating medium conveying means and cooling medium conveying means for conveying a medium to be heated and a medium to be cooled with respect to the side heat exchanger and the low pressure side heat exchanger, the high pressure side being operated at a supercritical pressure. In the used refrigeration cycle apparatus, the output shaft of the expansion mechanism section is connected to at least one drive shaft among the compressor, the heated medium conveying means, and the cooled medium conveying means.

この発明によれば、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結し、膨張機構部に連結する圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の駆動力の一部として、膨張機構部の回転力を利用するようにしている。   According to this invention, the output shaft of the expansion mechanism unit is connected to at least one drive shaft of the compressor, the heated medium transport unit, and the cooled medium transport unit, and the compression is coupled to the expansion mechanism unit. The rotational force of the expansion mechanism is used as part of the driving force of the machine, heated medium conveying means, and cooled medium conveying means.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記高圧側熱交換器の出口部を冷却する低温熱源をさらに備えたことを特徴とする。   The refrigeration cycle apparatus according to the next invention is characterized in that, in the above-mentioned invention, the refrigeration cycle apparatus further comprises a low-temperature heat source for cooling an outlet portion of the high-pressure side heat exchanger.

この発明によれば、高圧側熱交換器の出口部を冷媒回路と異なる低温熱源によって冷却し、放熱器出口温度を大幅に低下させるようにしている。   According to the present invention, the outlet portion of the high-pressure side heat exchanger is cooled by the low-temperature heat source different from that of the refrigerant circuit, and the radiator outlet temperature is greatly reduced.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、冷凍サイクルを行う第２冷凍サイクル手段をさらに備え、前記低温熱源は、前記第２冷凍サイクル手段の蒸発器であることを特徴とする。   The refrigeration cycle apparatus according to the next invention is characterized in that, in the above invention, the refrigeration cycle apparatus further includes second refrigeration cycle means for performing a refrigeration cycle, and the low-temperature heat source is an evaporator of the second refrigeration cycle means.

この発明によれば、高圧側熱交換器の出口部を、冷媒回路の冷凍サイクルとは異なる第２冷凍サイクル手段の蒸発器を用いて冷却し、放熱器出口温度を大幅に低下させるようにしている。   According to the present invention, the outlet portion of the high-pressure side heat exchanger is cooled by using the evaporator of the second refrigeration cycle means different from the refrigeration cycle of the refrigerant circuit, so that the radiator outlet temperature is greatly reduced. Yes.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記高圧側熱交換器および前記低圧側熱交換器を形成する伝熱管の管外径を５ｍｍ以下とすることを特徴とする。   The refrigeration cycle apparatus according to the next invention is characterized in that, in the above invention, the outer diameter of the heat transfer tube forming the high pressure side heat exchanger and the low pressure side heat exchanger is 5 mm or less.

この発明によれば、高圧側熱交換器および低圧側熱交換器を形成する伝熱管の管外径を５ｍｍ以下とし、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、冷媒質量速度を増加させ、十分な冷房および暖房能力を得るようにしている。 According to this invention, the pipe outer diameter of the heat transfer tubes forming the high-pressure side heat exchanger and the low-pressure side heat exchanger is set to 5 mm or less, and the characteristics of the CO 2 refrigerant as the supercritical pressure refrigerant are utilized to obtain the refrigerant mass rate. In order to obtain sufficient cooling and heating capacity.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記液配管および前記ガス配管を同一管径とすることを特徴とする。   The refrigeration cycle apparatus according to the next invention is characterized in that, in the above invention, the liquid pipe and the gas pipe have the same diameter.

この発明によれば、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、液配管およびガス配管を同一管径とし、液配管とガス配管との区別をなくすようにしている。 According to the present invention, the liquid pipe and the gas pipe have the same pipe diameter by utilizing the characteristics of the CO 2 refrigerant as the supercritical pressure refrigerant, and the distinction between the liquid pipe and the gas pipe is eliminated.

つぎの発明にかかる冷凍サイクル装置は、上記の発明において、前記液配管および前記ガス配管の長さを１５０ｍ以上としたことを特徴とする。   The refrigeration cycle apparatus according to the next invention is characterized in that, in the above invention, the length of the liquid pipe and the gas pipe is 150 m or more.

この発明によれば、ガス配管における圧力損失が小さいという、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、液配管およびガス配管の長さを１５０ｍ以上としている。 According to the present invention, the length of the liquid pipe and the gas pipe is set to 150 m or more by utilizing the characteristic of the CO 2 refrigerant as the supercritical pressure refrigerant that the pressure loss in the gas pipe is small.

つぎの発明にかかる冷凍サイクル装置の制御方法は、少なくとも圧縮機、高圧側熱交換器、絞り手段、および低圧側熱交換器を液配管およびガス配管を用いて接続した冷媒回路を形成するとともに、前記高圧側熱交換器および前記低圧側熱交換器に対して被加熱媒体および被冷却媒体を搬送する被加熱媒体搬送手段および被冷却媒体搬送手段を有し、高圧側が超臨界圧力で運転される冷媒を用いた冷凍サイクル装置の制御方法において、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持する保持工程と、前記高圧側熱交換器の出口配管温度を検出する検出工程と、前記検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御する制御工程とを含むことを特徴とする。   The control method of the refrigeration cycle apparatus according to the next invention forms a refrigerant circuit in which at least a compressor, a high-pressure side heat exchanger, a throttle means, and a low-pressure side heat exchanger are connected using a liquid pipe and a gas pipe, It has a heated medium conveying means and a cooled medium conveying means for conveying a medium to be heated and a medium to be cooled to the high pressure side heat exchanger and the low pressure side heat exchanger, and the high pressure side is operated at a supercritical pressure. In the method for controlling a refrigeration cycle apparatus using a refrigerant, a holding step for maintaining a relationship between an outlet pipe temperature of the high pressure side heat exchanger and a high pressure range in which a coefficient of performance of the refrigeration cycle is near the maximum, and the high pressure side heat Based on the relationship between the detection step of detecting the outlet pipe temperature of the exchanger, the detection result of the detection means and the holding means, the opening degree of the throttle means, the rotational speed of the compressor, and the heated Wherein the body of the flow rate, and the controls at least one of the flow rate of the cooling medium and a control step of controlling the high pressure range.

この発明によれば、保持工程によって、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持し、検出工程によって、前記高圧側熱交換器の出口配管温度を検出し、制御工程によって、前記検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしている。   According to the present invention, the holding step maintains the relationship between the outlet pipe temperature of the high pressure side heat exchanger and the high pressure range where the coefficient of performance of the refrigeration cycle is near the maximum, and the high pressure side heat exchange is detected by the detecting step. And detecting the opening temperature of the throttle means, the rotation speed of the compressor, and the flow rate of the heated medium based on the detection result of the detection means and the relationship between the holding means. And at least one of the flow rates of the medium to be cooled is controlled within the high pressure range.

つぎの発明にかかる冷凍サイクル装置の制御方法は、上記の発明において、前記検出工程は、前記圧縮機から前記高圧側交換器に至る配管の温度を検出する配管温度検出工程をさらに含み、前記制御工程は、前記配管温度検出工程によって検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御することを特徴とする。   In the control method of the refrigeration cycle apparatus according to the next invention, in the above invention, the detection step further includes a pipe temperature detection step of detecting a temperature of a pipe from the compressor to the high-pressure side exchanger, The process estimates the high pressure based on the temperature detected by the pipe temperature detection process, and based on the estimated high pressure, the detection result of the detection means, and the relationship between the holding means, Controlling at least one of the opening degree, the rotation speed of the compressor, the flow rate of the heated medium, and the flow rate of the cooled medium to control the high pressure range.

この発明によれば、前記検出工程の配管温度検出工程によって、前記圧縮機から前記高圧側交換器に至る配管の温度をさらに検出し、前記制御工程によって、前記配管温度検出工程によって検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしている。   According to this invention, the temperature of the pipe from the compressor to the high-pressure side exchanger is further detected by the pipe temperature detection process of the detection process, and the temperature detected by the pipe temperature detection process is detected by the control process. Based on the estimated high pressure, the detection result of the detection means, and the relationship of the holding means, the opening degree of the throttle means, the rotation speed of the compressor, and the heated At least one of the flow rate of the medium and the flow rate of the medium to be cooled is controlled to be within the high pressure range.

この発明によれば、保持手段が、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持し、制御手段が、前記高圧側熱交換器の出口配管温度を検出する検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにし、被加熱媒体あるいは被冷却媒体の流量が変化した場合であっても高圧圧力範囲で運転できるようにしているので、常に成績係数ＣＯＰの高い運転を行うことができ、省エネルギー化を達成する冷凍サイクル装置を実現することができるという効果を奏する。   According to this invention, the holding means holds the relationship between the outlet pipe temperature of the high pressure side heat exchanger and the high pressure range where the coefficient of performance of the refrigeration cycle is near the maximum, and the control means is the high pressure side heat exchange. Based on the detection result of the detection means for detecting the outlet pipe temperature of the compressor and the relationship between the holding means, the opening degree of the throttle means, the rotation speed of the compressor, the flow rate of the heated medium, and the cooled medium Since at least one of the flow rates is controlled to be within the high pressure range, and even when the flow rate of the heating medium or the cooling medium changes, it can be operated in the high pressure range. It is possible to always perform an operation with a high coefficient of performance COP and to realize an refrigeration cycle apparatus that achieves energy saving.

つぎの発明によれば、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結し、膨張機構部に連結する圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の駆動力の一部として、膨張機構部の回転力を利用するようにしているので、圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の必要駆動力を低減することができるとともに、常に高圧圧力範囲で運転できるようにしているため、常に成績係数ＣＯＰの高い運転を行うことができ、省エネルギー化を達成する冷凍サイクル装置を実現することができるという効果を奏する。   According to the next invention, the output shaft of the expansion mechanism section is connected to at least one drive shaft of the compressor, the heated medium transfer means, and the cooled medium transfer means, and is connected to the expansion mechanism section. Since the rotational force of the expansion mechanism is used as part of the driving force of the compressor, heated medium conveying means, and cooled medium conveying means, the compressor, heated medium conveying means, cooled medium conveyance Since the required driving force of the means can be reduced and the operation can always be performed in the high pressure range, the operation with a high coefficient of performance COP can always be performed, and a refrigeration cycle apparatus that achieves energy saving is realized. There is an effect that can be.

つぎの発明によれば、制御手段が、前記配管温度検出手段が検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしているので、圧力を直接測定する高価な圧力検出器を用いる必要がなく、簡易な構成によって安価な冷凍サイクル装置を実現することができるという効果を奏する。   According to the next invention, the control means estimates the high pressure based on the temperature detected by the pipe temperature detecting means, and the relationship between the estimated high pressure, the detection result of the detecting means, and the holding means is determined. Basically, at least one of the opening degree of the throttle means, the rotation speed of the compressor, the flow rate of the heated medium, and the flow rate of the cooled medium is controlled to control the high pressure range. Therefore, it is not necessary to use an expensive pressure detector that directly measures the pressure, and an inexpensive refrigeration cycle apparatus can be realized with a simple configuration.

つぎの発明によれば、前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結し、膨張機構部に連結する圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の駆動力の一部として、膨張機構部の回転力を利用するようにしているので、圧縮機、被加熱媒体搬送手段、被冷却媒体搬送手段の必要駆動力を低減することができ、省エネルギー化を達成する冷凍サイクル装置を実現することができるという効果を奏する。   According to the next invention, the output shaft of the expansion mechanism section is connected to at least one drive shaft of the compressor, the heated medium transfer means, and the cooled medium transfer means, and is connected to the expansion mechanism section. Since the rotational force of the expansion mechanism is used as part of the driving force of the compressor, heated medium conveying means, and cooled medium conveying means, the compressor, heated medium conveying means, cooled medium conveyance The required driving force of the means can be reduced, and an effect is achieved that a refrigeration cycle apparatus that achieves energy saving can be realized.

つぎの発明によれば、高圧側熱交換器の出口部を冷媒回路と異なる低温熱源によって冷却し、放熱器出口温度を大幅に低下させるようにしているので、冷媒回路の冷凍サイクルと低温熱源とを含めたシステム全体としての省エネルギー化を図ることができるという効果を奏する。   According to the next invention, the outlet portion of the high-pressure side heat exchanger is cooled by a low-temperature heat source different from the refrigerant circuit, and the radiator outlet temperature is greatly reduced. There is an effect that energy saving as a whole system including can be achieved.

つぎの発明によれば、高圧側熱交換器の出口部を、冷媒回路の冷凍サイクルとは異なる第２冷凍サイクル手段の蒸発器を用いて冷却し、放熱器出口温度を大幅に低下させるようにしているので、冷媒回路の冷凍サイクルと第２冷凍サイクル手段の冷凍サイクルとを含めたシステム全体としての省エネルギー化を図ることができるという効果を奏する。   According to the next invention, the outlet portion of the high-pressure side heat exchanger is cooled using the evaporator of the second refrigeration cycle means different from the refrigeration cycle of the refrigerant circuit, so that the radiator outlet temperature is greatly reduced. Therefore, there is an effect that energy saving can be achieved as the entire system including the refrigeration cycle of the refrigerant circuit and the refrigeration cycle of the second refrigeration cycle means.

つぎの発明によれば、高圧側熱交換器および低圧側熱交換器を形成する伝熱管の管外径を５ｍｍ以下とし、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、冷媒質量速度を増加させ、十分な冷房および暖房能力を得るようにしているので、超臨界冷凍サイクルに適したＣＯ2冷媒に対応した冷凍サイクル装置の熱交換器を実現することができるとともに、熱交換器の小型軽量化を促進することができるという効果を奏する。 According to the next invention, the outer diameter of the heat transfer tube forming the high-pressure side heat exchanger and the low-pressure side heat exchanger is set to 5 mm or less, and the mass of the refrigerant is obtained by utilizing the characteristics of the CO 2 refrigerant as the supercritical pressure refrigerant. Since the speed is increased and sufficient cooling and heating capacity is obtained, a heat exchanger for a refrigeration cycle apparatus compatible with a CO 2 refrigerant suitable for a supercritical refrigeration cycle can be realized, and a heat exchanger There is an effect that the reduction in size and weight can be promoted.

つぎの発明によれば、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、液配管およびガス配管を同一管径とし、液配管とガス配管との区別をなくすようにしているので、冷媒配管工事にかかる作業性を向上させることができるという効果を奏する。 According to the next invention, using the characteristics of the CO 2 refrigerant as the supercritical pressure refrigerant, the liquid pipe and the gas pipe have the same diameter, so that the distinction between the liquid pipe and the gas pipe is eliminated. There exists an effect that workability | operativity concerning refrigerant | coolant piping construction can be improved.

つぎの発明によれば、ガス配管における圧力損失が小さいという、超臨界圧力冷媒としてのＣＯ2冷媒の特性を利用して、液配管およびガス配管の長さを１５０ｍ以上としているので、性能低下が十分許容し得る大規模ビルなどの空気調和システムに用いることができるという効果を奏する。 According to the next invention, the length of the liquid pipe and the gas pipe is set to 150 m or more by utilizing the characteristic of the CO 2 refrigerant as the supercritical pressure refrigerant that the pressure loss in the gas pipe is small. There is an effect that it can be used for an air-conditioning system such as a large-scale building that is sufficiently acceptable.

つぎの発明によれば、保持工程によって、前記高圧側熱交換器の出口配管温度と冷凍サイクルの成績係数が最大付近となる高圧圧力範囲との関係を保持し、検出工程によって、前記高圧側熱交換器の出口配管温度を検出し、制御工程によって、前記検出手段の検出結果および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしているので、常に成績係数ＣＯＰの高い運転を行うことができ、省エネルギー化を達成する冷凍サイクル装置を実現することができるという効果を奏する。   According to the next invention, the holding step maintains the relationship between the outlet pipe temperature of the high-pressure side heat exchanger and the high-pressure range where the coefficient of performance of the refrigeration cycle is near the maximum, and the high-pressure side heat is detected by the detecting step. The outlet pipe temperature of the exchanger is detected, and the control step determines the opening degree of the throttle means, the rotation speed of the compressor, the heating medium based on the detection result of the detection means and the relationship between the holding means. Since at least one of the flow rate and the flow rate of the medium to be cooled is controlled to be within the high pressure range, the operation with a high coefficient of performance COP can always be performed, and refrigeration that achieves energy saving There exists an effect that a cycle device can be realized.

つぎの発明によれば、前記検出工程の配管温度検出工程によって、前記圧縮機から前記高圧側交換器に至る配管の温度をさらに検出し、前記制御工程によって、前記配管温度検出工程によって検出した温度をもとに高圧圧力を推定し、この推定した高圧圧力、前記検出手段の検出結果、および前記保持手段の関係をもとに、前記絞り手段の開度、前記圧縮機の回転数、前記被加熱媒体の流量、および前記被冷却媒体の流量のうち少なくとも１つを制御して前記高圧圧力範囲に制御するようにしているので、圧力を直接測定する高価な圧力検出器を用いる必要がなく、安価な冷凍サイクル装置を実現することができるという効果を奏する。   According to the next invention, the pipe temperature detection step of the detection step further detects the temperature of the pipe from the compressor to the high-pressure side exchanger, and the control step detects the temperature detected by the pipe temperature detection step. Based on the estimated high pressure, the detection result of the detection means, and the relationship of the holding means, the opening of the throttle means, the rotational speed of the compressor, Since at least one of the flow rate of the heating medium and the flow rate of the medium to be cooled is controlled to be within the high pressure range, there is no need to use an expensive pressure detector that directly measures the pressure, There is an effect that an inexpensive refrigeration cycle apparatus can be realized.

以下に添付図面を参照して、この発明にかかる冷凍サイクル装置の好適な実施の形態を詳細に説明する。   Exemplary embodiments of a refrigeration cycle apparatus according to the present invention will be explained below in detail with reference to the accompanying drawings.

実施の形態１．
図１は、この発明の実施の形態１である冷凍サイクル装置の構成を示す図である。図１において、この冷凍サイクル装置は、空気調和機であり、冷媒として、高圧側が臨界圧力以上となる二酸化炭素のような超臨界冷媒を用いている。この空気調和機は、図１５に示した空気調和機と同様に、室内熱交換器６、室内ファン７を含む室内ユニットＡと、室外熱交換器２、室外ファン３、圧縮機１およびアキュムレータ１２を含む室外ユニットＢとを有し、室内ユニットＡと室外ユニットＢとは、接続配管（液配管）５および接続配管（ガス配管）８によって接続されている。なお、図１に示した空気調和機では、室内ユニットＡが１台だけ接続された場合を示しているが、１台の室外ユニットＢに対して、室内ユニットＡが複数台設置される構成としても良い。なお、図１５に示した空気調和機と同一構成部分には、同一符号を付している。 Embodiment 1 FIG.
1 is a diagram showing a configuration of a refrigeration cycle apparatus according to Embodiment 1 of the present invention. In FIG. 1, this refrigeration cycle apparatus is an air conditioner, and uses a supercritical refrigerant such as carbon dioxide whose high pressure side is equal to or higher than the critical pressure as a refrigerant. This air conditioner includes an indoor unit A including an indoor heat exchanger 6 and an indoor fan 7, an outdoor heat exchanger 2, an outdoor fan 3, a compressor 1, and an accumulator 12, similarly to the air conditioner shown in FIG. 15. The indoor unit A and the outdoor unit B are connected by a connection pipe (liquid pipe) 5 and a connection pipe (gas pipe) 8. In the air conditioner shown in FIG. 1, only one indoor unit A is connected. However, a plurality of indoor units A are installed for one outdoor unit B. Also good. In addition, the same code | symbol is attached | subjected to the same component as the air conditioner shown in FIG.

室内ファン７は、室内の空気を室内熱交換器６に導入させ、室外ファン３は、室外の空気を室外熱交換器２に導入させる。圧縮機１は、四方切替弁９を介して、室内熱交換器６および室外熱交換器２に接続され、室内熱交換器６および室外熱交換器２を接続する他方の液配管５には、絞り手段としての電子式膨張弁４が設けられる。   The indoor fan 7 introduces indoor air into the indoor heat exchanger 6, and the outdoor fan 3 introduces outdoor air into the outdoor heat exchanger 2. The compressor 1 is connected to the indoor heat exchanger 6 and the outdoor heat exchanger 2 via the four-way switching valve 9, and the other liquid pipe 5 connecting the indoor heat exchanger 6 and the outdoor heat exchanger 2 includes An electronic expansion valve 4 is provided as a throttle means.

圧縮機１は、吸入ガスを圧縮して室外熱交換器２または室内熱交換器６に送る。アキュムレータ１２は、圧縮機１の吸入側に接続され、室外熱交換器２または室内熱交換器６から戻る液冷媒を貯留する。このアキュムレータ１２の上部には、吸入配管１３ｂの一端が接続され、他端が圧縮機１に接続され、アキュムレータ１２内で気液分離した冷媒ガスが吸入配管１３ｂを通って圧縮機１に供給される。さらに、アキュムレータ１２の上部には、室外熱交換器２（暖房運転時）または室内熱交換器６（冷房運転時）からの冷媒を、四方切替弁９を介してアキュムレータ１２に流入させる流入管１３ａが接続される。また、圧縮機１の吐出側配管には、冷媒の吐出温度を検出する配管温度検出器２２が設けられ、室内熱交換器６の出入口には、配管温度検出器２０，２１が設けられ、室外熱交換器２の冷房時の出口（暖房時の入口）には、配管温度検出器２３が設けられる。   The compressor 1 compresses the suction gas and sends it to the outdoor heat exchanger 2 or the indoor heat exchanger 6. The accumulator 12 is connected to the suction side of the compressor 1 and stores liquid refrigerant returning from the outdoor heat exchanger 2 or the indoor heat exchanger 6. One end of the suction pipe 13b is connected to the upper portion of the accumulator 12, the other end is connected to the compressor 1, and the refrigerant gas separated in the accumulator 12 is supplied to the compressor 1 through the suction pipe 13b. The Further, an inflow pipe 13a through which refrigerant from the outdoor heat exchanger 2 (during heating operation) or the indoor heat exchanger 6 (during cooling operation) flows into the accumulator 12 via the four-way switching valve 9 is disposed above the accumulator 12. Is connected. The discharge side pipe of the compressor 1 is provided with a pipe temperature detector 22 that detects the refrigerant discharge temperature, and the indoor heat exchanger 6 is provided with pipe temperature detectors 20 and 21 at the inlet and outlet. A piping temperature detector 23 is provided at the cooling outlet of the heat exchanger 2 (heating inlet).

ここで、この冷凍サイクル装置の動作について説明すると、まず、冷房運転時において、四方切替弁９は、図１の実線で示すように接続され、圧縮機１と室外熱交換器２とが連通し、室内熱交換器６および室外熱交換器２は、それぞれ蒸発器および放熱器として機能する。すなわち、圧縮機１から吐出された高温・高圧の冷媒ガスが室外熱交換器２に導入され、ここで、冷媒ガスと室外空気との熱交換が行われた後、中温・高圧ガスは電子式膨張弁４で減圧されて低温・低圧の二相冷媒となり、液配管５を通って室内熱交換器６に導入され、室内空気と熱交換が行われた後、再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→室外熱交換器２→電子式膨張弁４→液配管５→室内熱交換器１→ガス配管８→四方切替弁９→アキュムレータ１２→圧縮機１の順で冷媒回路内を流動し、冷房運転が行われる。この時、コントローラ１０は、冷房能力の制御を、配管温度検出器２１で検出された蒸発器入口温度ＥＴが目標蒸発温度ＥＴ*となるように、圧縮機１の回転数を制御することによって行う。   Here, the operation of this refrigeration cycle apparatus will be described. First, during the cooling operation, the four-way switching valve 9 is connected as shown by the solid line in FIG. 1 so that the compressor 1 and the outdoor heat exchanger 2 communicate with each other. The indoor heat exchanger 6 and the outdoor heat exchanger 2 function as an evaporator and a radiator, respectively. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the outdoor heat exchanger 2, where after the heat exchange between the refrigerant gas and the outdoor air is performed, the medium-temperature and high-pressure gas is electronic. The refrigerant is decompressed by the expansion valve 4 to become a low-temperature / low-pressure two-phase refrigerant, introduced into the indoor heat exchanger 6 through the liquid pipe 5, exchanged with indoor air, and then sucked into the compressor 1 again. . That is, the refrigerant is compressor 1 → four-way switching valve 9 → outdoor heat exchanger 2 → electronic expansion valve 4 → liquid piping 5 → indoor heat exchanger 1 → gas piping 8 → four-way switching valve 9 → accumulator 12 → compressor. The refrigerant flows in the order of 1, and the cooling operation is performed. At this time, the controller 10 controls the cooling capacity by controlling the rotation speed of the compressor 1 so that the evaporator inlet temperature ET detected by the pipe temperature detector 21 becomes the target evaporation temperature ET *. .

一方、暖房運転時において、四方切替弁９は、図１の点線で示すように接続され、圧縮機１と室内熱交換器６とが連通し、室内熱交換器６および室外熱交換器２はそれぞれ放熱器および蒸発器として機能する。すなわち、圧縮機１より吐出された高温・高圧の冷媒ガスが室内熱交換器６に導入され、ここで、冷媒ガスと室外空気との熱交換が行われた後、中温・高圧ガスは液配管５を通過した後、電子式膨張弁４で減圧され、室外熱交換器２に導入されて室外空気と熱交換が行われ、再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→室内熱交換器６→液配管５→電子式膨張弁４→室外熱交換器２→ガス配管３→四方切替弁９→アキュムレータ１２→圧縮機１の順で冷媒回路内を流動し、暖房運転が行われる。この時、コントローラ１０は、暖房能力の制御を、配管温度検出器２２で検出された吐出温度Ｔｄを用いて、高圧圧力ＨＰを算出し、このＨＰが目標高圧圧力ＨＰ１*となるように圧縮機１の回転数を制御することによって行う。   On the other hand, during the heating operation, the four-way switching valve 9 is connected as shown by the dotted line in FIG. 1, the compressor 1 and the indoor heat exchanger 6 communicate with each other, and the indoor heat exchanger 6 and the outdoor heat exchanger 2 are Each functions as a radiator and an evaporator. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the indoor heat exchanger 6 where heat exchange between the refrigerant gas and the outdoor air is performed, and then the intermediate-temperature and high-pressure gas is liquid pipe. After passing through 5, the pressure is reduced by the electronic expansion valve 4, introduced into the outdoor heat exchanger 2, heat exchange with the outdoor air is performed, and again sucked into the compressor 1. That is, the refrigerant is compressor 1 → four-way switching valve 9 → indoor heat exchanger 6 → liquid piping 5 → electronic expansion valve 4 → outdoor heat exchanger 2 → gas piping 3 → four-way switching valve 9 → accumulator 12 → compressor. Heating operation is performed by flowing in the refrigerant circuit in the order of 1. At this time, the controller 10 calculates the high pressure HP using the discharge temperature Td detected by the pipe temperature detector 22 and controls the heating capacity so that the HP becomes the target high pressure HP1 *. This is done by controlling the number of revolutions of 1.

さらに、コントローラ１０は、室外ファン３および室内ファン７の各モータ３ａ，７ａの回転数をそれぞれ制御するとともに、配管温度検出器２０〜２３で検出された各検出値から、電子式膨張弁４の開度や圧縮機１の回転数を制御する。ここで、コントローラ１０には、メモリ１０ａが備えられ、予め、高圧側熱交換器（冷房時は室外熱交換器２、暖房時は室内熱交換器６）の出口温度すなわち放熱器出口温度と、成績係数ＣＯＰが最大付近となる最適高圧圧力ＨＰ*との特性が記憶されている。   Further, the controller 10 controls the rotational speeds of the motors 3a and 7a of the outdoor fan 3 and the indoor fan 7, respectively, and from the detected values detected by the pipe temperature detectors 20 to 23, the electronic expansion valve 4 is controlled. The opening degree and the rotation speed of the compressor 1 are controlled. Here, the controller 10 is provided with a memory 10a, and the outlet temperature of the high-pressure side heat exchanger (the outdoor heat exchanger 2 during cooling and the indoor heat exchanger 6 during heating), that is, the radiator outlet temperature, The characteristic with the optimum high pressure HP * at which the coefficient of performance COP is near the maximum is stored.

ここで、放熱器出口温度に対して成績係数ＣＯＰが最大付近となる最適な高圧圧力が存在する理由について、図２および図３を参照して説明する。図２は、超臨界冷凍サイクルの圧力−エンタルピー線図である。図２に示した等温線(３５℃、４５℃、５５℃)のうち、たとえば、３５℃の等温線が、放熱器出口状態を示している。高圧圧力を、圧力ＰＡと圧力ＰＢとの間の圧力Ｍ0に設定して、超臨界冷凍サイクルを運転していた場合、たとえば冷房時の成績係数ＣＯＰは、(ｈｅO−ｈｅM)／(ｈｃM−ｈｅO)となる。 Here, the reason why there is an optimum high pressure at which the coefficient of performance COP is near the maximum with respect to the radiator outlet temperature will be described with reference to FIGS. 2 and 3. FIG. 2 is a pressure-enthalpy diagram of the supercritical refrigeration cycle. Among the isotherms shown in FIG. 2 (35 ° C., 45 ° C., 55 ° C.), for example, an isotherm of 35 ° C. indicates a radiator outlet state. When the supercritical refrigeration cycle is operated with the high pressure set to the pressure M 0 between the pressure PA and the pressure PB, for example, the coefficient of performance COP during cooling is (he O −he M ) / ( hc M −he O ).

ここで、高圧圧力を圧力Ｍ0から圧力ＰＡに上昇させると、蒸発器内のエンタルピー差は(ｈｅM−ｈｅA)増加するのに対して、圧縮機入力が(ｈｃA−ｈｃM)増加する。ここで、エンタルピー差の増加量（ｈｅM−ｈｅA）は、入力の増加量（ｈｃA−ｈｃM）に比して小さく、結果として成績係数ＣＯＰは低下する。逆に、高圧圧力を圧力Ｍ0から圧力ＰＢに低下させると、エンタルピー差が（ｈｅB−ｈｅM）減少するのに対して、入力が（ｈｃM−ｈｃB）減少する。エンタルピー差の減少量（ｈｅB−ｈｅM）は、入力の減少量（ｈｃM−ｈｃB）に比して大きく、同様に成績係数ＣＯＰは低下する。このように、成績係数ＣＯＰの最大値を与える高圧圧力は、等温線の曲線形状によって一意的に決定され、放熱器出口温度が与えられた場合に、最適高圧圧力を示す点の軌跡は点線Ｌ１のようになる。 Here, when the high pressure is increased from the pressure M 0 to the pressure PA, the enthalpy difference in the evaporator increases (he M −he A ), whereas the compressor input increases (hc A −hc M ). To do. Here, the amount of increase in enthalpy difference (he M −he A ) is smaller than the amount of increase in input (hc A −hc M ), and as a result, the coefficient of performance COP decreases. Conversely, when the high pressure is decreased from the pressure M 0 to the pressure PB, the enthalpy difference decreases (he B −he M ), whereas the input decreases (hc M −hc B ). The amount of decrease in enthalpy difference (he B −he M ) is larger than the amount of decrease in input (hc M −hc B ), and similarly, the coefficient of performance COP decreases. As described above, the high pressure that gives the maximum value of the coefficient of performance COP is uniquely determined by the curve shape of the isotherm, and when the radiator outlet temperature is given, the locus of the point indicating the optimum high pressure is the dotted line L1. become that way.

図３は、上述した関係を、放熱器出口温度をパラメータとして、横軸に高圧圧力をとり、縦軸に成績係数ＣＯＰをとった場合を示す図である。図３において、成績係数ＣＯＰが最大値を示す高圧圧力（図３中ではＭ0，Ｍ1，Ｍ2）と、放熱器出口温度との関係を整理すると、曲線Ｌ２で示される特性を得ることができる。したがって、この曲線Ｌ２の最適関係をもとに、放熱器出口温度毎に、成績係数ＣＯＰが最大付近となる高圧圧力範囲を定め、冷凍サイクルの運転を制御することによって、高効率運転が可能となるのである。 FIG. 3 is a diagram showing the above-described relationship, where the horizontal axis represents high pressure and the vertical axis represents the coefficient of performance COP, with the radiator outlet temperature as a parameter. In FIG. 3, when the relationship between the high pressure at which the coefficient of performance COP has the maximum value (M 0 , M 1 , M 2 in FIG. 3) and the radiator outlet temperature is arranged, the characteristic indicated by the curve L2 is obtained. Can do. Therefore, based on the optimum relationship of the curve L2, for each radiator outlet temperature, a high pressure range in which the coefficient of performance COP is near the maximum is determined, and the operation of the refrigeration cycle is controlled. It becomes.

つぎに、図４に示すフローチャートを参照して、冷房運転時におけるコントローラ１０の制御処理手順について説明する。この冷房運転の制御処理では、冷房能力の確保に必要な蒸発温度を圧縮機１の回転数によって制御し、蒸発器出口状態を電子式膨張弁４の開度によって制御し、高圧圧力を室外ファン３の回転数によって制御するものである。ここで、予め、コントローラ１０のメモリ１０ａには、上述したように、放熱器出口温度ＣＴ０と冷凍サイクルの成績係数ＣＯＰが最大付近となる目標高圧圧力ＨＰ*との特性式が設定されている。   Next, the control processing procedure of the controller 10 during the cooling operation will be described with reference to the flowchart shown in FIG. In the control process of the cooling operation, the evaporation temperature necessary for securing the cooling capacity is controlled by the rotation speed of the compressor 1, the evaporator outlet state is controlled by the opening degree of the electronic expansion valve 4, and the high pressure is controlled by the outdoor fan. 3 is controlled by the number of rotations. Here, as described above, the characteristic equation between the radiator outlet temperature CT0 and the target high pressure HP * at which the coefficient of performance COP of the refrigeration cycle is near the maximum is set in the memory 10a of the controller 10 in advance.

図４において、まず、コントローラ１０は、配管温度検出器２０〜２３から検出値ＥＴ０，ＥＴ，Ｔｄ，ＣＴ０を取得する（ステップＳ１）。その後、コントローラ１０は、蒸発温度ＥＴと目標蒸発温度ＥＴ*との偏差ΔＥＴを算出し（ステップＳ２）、さらに、蒸発器出口過熱度ＳＨと目標過熱度ＳＨ*との偏差ΔＳＨを算出する（ステップＳ３）。さらに、コントローラ１０は、メモリ１０ａから放熱器出口温度ＣＴ０に対応した目標高圧圧力ＨＰ*を算出する（ステップＳ４）。その後、コントローラ１０は、配管温度検出器２２からの検出値を用いて演算した高圧圧力ＨＰと目標高圧圧力ＨＰ*との偏差ΔＨＰを求める（ステップＳ５）。   In FIG. 4, first, the controller 10 acquires detection values ET0, ET, Td, and CT0 from the pipe temperature detectors 20 to 23 (step S1). Thereafter, the controller 10 calculates a deviation ΔET between the evaporation temperature ET and the target evaporation temperature ET * (step S2), and further calculates a deviation ΔSH between the evaporator outlet superheat degree SH and the target superheat degree SH * (step S2). S3). Further, the controller 10 calculates a target high pressure HP * corresponding to the radiator outlet temperature CT0 from the memory 10a (step S4). Thereafter, the controller 10 obtains a deviation ΔHP between the high pressure HP calculated using the detected value from the pipe temperature detector 22 and the target high pressure HP * (step S5).

ここで、蒸発温度ＥＴは、配管温度検出器２１からの検出値であり、目標蒸発温度ＥＴ*は、たとえば運転台数や外気温度の関数として定められる。また、蒸発器出口過熱度ＳＨは、蒸発器出口温度ＥＴ０と蒸発温度ＥＴとの差（すなわち、ＳＨ＝ＥＴ０−ＥＴ）であり、目標過熱度ＳＨ*は、必要な冷房能力が確保できるように、たとえば１〜１０℃の値が設定される。   Here, the evaporation temperature ET is a detected value from the pipe temperature detector 21, and the target evaporation temperature ET * is determined as a function of, for example, the number of operating units and the outside air temperature. Further, the evaporator outlet superheat degree SH is a difference between the evaporator outlet temperature ET0 and the evaporation temperature ET (ie, SH = ET0−ET), and the target superheat degree SH * can ensure the necessary cooling capacity. For example, a value of 1 to 10 ° C. is set.

ここで、配管温度検出器２２の検出値を用いた高圧圧力ＨＰの求め方について図５を参照して説明する。一般に、圧縮機１の吐出温度Ｔｄは、高圧圧力ＨＰ、低圧圧力ＬＰ、飽和蒸気エンタルピーｈｅOおよび図示しない圧縮機回転数のデータと、予め実験などによって求められた圧縮機１の性能曲線から断熱効率（図５における直線Ｌ３の傾きに相当）を算出して推定することができる。 Here, how to obtain the high pressure HP using the detection value of the pipe temperature detector 22 will be described with reference to FIG. In general, the discharge temperature Td of the compressor 1 is adiabatic from the data of the high pressure HP, the low pressure LP, the saturated steam enthalpy he O, the compressor rotation speed (not shown), and the performance curve of the compressor 1 obtained in advance through experiments or the like. Efficiency (corresponding to the slope of the straight line L3 in FIG. 5) can be calculated and estimated.

したがって、逆に、圧縮機１の吐出温度Ｔｄ、蒸発温度ＥＴ、圧縮機回転数の情報を得ることができれば、冷媒の物性式を用いて蒸発温度ＥＴから低圧圧力ＬＰと飽和蒸気エンタルピーｈｅOをそれぞれ算出することができ、さらに圧縮機１の性能曲線を用いて高圧圧力ＨＰを推定することができる。そして、これらの関係式は、予め、コントローラ１０内のメモリ１０ａに記憶させておく。実際には、電子式膨張弁４の開度によって、圧縮機１の吸入エンタルピーが飽和蒸気エンタルピーと異なる場合が生じ（図５におけるｈｅO'）、断熱効率が飽和蒸気を吸入した場合と異なる場合が生じるため（図５における直線Ｌ４の傾きに相当）、電子式膨張弁４の開度によって高圧圧力ＨＰの推定値を補正するように構成しても良い。 Therefore, conversely, if information on the discharge temperature Td, the evaporation temperature ET, and the compressor rotation speed of the compressor 1 can be obtained, the low-pressure pressure LP and the saturated vapor enthalpy he O are calculated from the evaporation temperature ET using the physical properties of the refrigerant. Each can be calculated, and the high pressure HP can be estimated using the performance curve of the compressor 1. These relational expressions are stored in the memory 10a in the controller 10 in advance. Actually, depending on the opening degree of the electronic expansion valve 4, the suction enthalpy of the compressor 1 may be different from the saturated steam enthalpy (he O ′ in FIG. 5), and the adiabatic efficiency is different from the case where the saturated steam is sucked. (Corresponding to the slope of the straight line L4 in FIG. 5), the estimated value of the high pressure HP may be corrected by the opening of the electronic expansion valve 4.

その後、コントローラ１０は、蒸発温度ＥＴの制御偏差ΔＥＴの絶対値｜ΔＥＴ｜が所定値ε１未満であるか否かを判断し（ステップＳ６）、所定値ε１未満でない場合（ステップＳ６，Ｎｏ）、冷房能力が不足または過剰であると判断し、圧縮機１の回転数を変更し（ステップＳ７）、ステップＳ８に移行する。一方、所定値ε１未満である場合（ステップＳ６，Ｙｅｓ)には、そのままステップＳ８に移行する。   Thereafter, the controller 10 determines whether or not the absolute value | ΔET | of the control deviation ΔET of the evaporation temperature ET is less than the predetermined value ε1 (step S6), and if not less than the predetermined value ε1 (step S6, No), It is determined that the cooling capacity is insufficient or excessive, the rotational speed of the compressor 1 is changed (step S7), and the process proceeds to step S8. On the other hand, if it is less than the predetermined value ε1 (step S6, Yes), the process proceeds to step S8 as it is.

ここで、所定値ε１は、制御を実施する最低偏差を示しており、たとえば０．５℃である。圧縮機回転数の変更方法は、たとえばΔＥＴ＞＋ε１の場合には、蒸発温度ＥＴが目標蒸発温度ＥＴ*に比して高いため、冷房能力が不足していると判断し、圧縮機１の回転数を増加させる。一方、ΔＥＴ＜−ε１の場合には、蒸発温度ＥＴが目標蒸発温度ＥＴ*に比して低いため、冷房能力が過剰であると判断し、圧縮機１の回転数を減少させる。圧縮機回転数の増加や減少割合は、制御偏差ΔＥＴの絶対値｜ΔＥＴ｜に応じて、たとえば周波数の変更幅ΔＨｚが、
ΔＨｚ＝α｜ΔＥＴ｜ （αは定数）
となる関係を満足するように変更すればよい。 Here, the predetermined value ε1 indicates the minimum deviation for performing the control, and is 0.5 ° C., for example. For example, when ΔET> + ε1, the compressor rotation speed is changed by determining that the cooling capacity is insufficient because the evaporation temperature ET is higher than the target evaporation temperature ET *. Increase the number. On the other hand, in the case of ΔET <−ε1, since the evaporation temperature ET is lower than the target evaporation temperature ET *, it is determined that the cooling capacity is excessive, and the rotational speed of the compressor 1 is decreased. The increase or decrease rate of the compressor rotational speed is, for example, a frequency change width ΔHz according to the absolute value | ΔET | of the control deviation ΔET,
ΔHz = α | ΔET | (α is a constant)
It may be changed so as to satisfy the relationship.

その後、コントローラ１０は、蒸発器出口過熱度ＳＨの制御偏差ΔＳＨの絶対値｜ΔＳＨ｜が所定値ε２未満であるか否かを判断する（ステップＳ８）。絶対値｜ΔＳＨ｜が所定値ε２未満でない場合（ステップＳ８，Ｎｏ）には、電子式膨張弁４の開度が過小または過大であると判断し、電子式膨張弁４の開度を変更し(ステップＳ９)、ステップＳ１０に移行する。一方、絶対値｜ΔＳＨ｜が所定値ε２未満である場合（ステップＳ８，Ｙｅｓ）には、そのままステップＳ１０に移行する。   Thereafter, the controller 10 determines whether or not the absolute value | ΔSH | of the control deviation ΔSH of the evaporator outlet superheat degree SH is less than a predetermined value ε2 (step S8). If the absolute value | ΔSH | is not less than the predetermined value ε2 (step S8, No), it is determined that the opening degree of the electronic expansion valve 4 is too small or too large, and the opening degree of the electronic expansion valve 4 is changed. (Step S9), the process proceeds to Step S10. On the other hand, if the absolute value | ΔSH | is less than the predetermined value ε2 (step S8, Yes), the process proceeds to step S10 as it is.

ここで、所定値ε２は、制御を実施する最低偏差を示しており、たとえば０．５℃である。電子式膨張弁４の開度の変更方法は、たとえばΔＳＨ＞＋ε２の場合には、蒸発器出口過熱度ＳＨが目標過熱度ＳＨ*に比して大きいため、電子式膨張弁４の開度が過小であると判断し、電子式膨張弁４の開度を増加させる。一方、ΔＳＨ＜−ε２の場合には、蒸発器出口過熱度ＳＨが目標過熱度ＳＨ*に比して低いため、電子式膨張弁４の開度が過大であると判断し、電子式膨張弁４の開度を減少させる。電子式膨張弁４の開度の増加や減少の割合は、圧縮機回転数の場合と同様に、制御偏差の絶対値｜ΔＳＨ｜に応じて変更するようにすればよい。   Here, the predetermined value ε2 indicates the minimum deviation for performing the control, and is 0.5 ° C., for example. For example, when ΔSH> + ε2, the degree of opening of the electronic expansion valve 4 is larger than the target degree of superheating SH *. It judges that it is too small, and increases the opening degree of the electronic expansion valve 4. On the other hand, when ΔSH <−ε2, the evaporator outlet superheat degree SH is lower than the target superheat degree SH *, so it is determined that the opening degree of the electronic expansion valve 4 is excessive, and the electronic expansion valve 4 is reduced. The rate of increase or decrease of the opening degree of the electronic expansion valve 4 may be changed according to the absolute value | ΔSH | of the control deviation as in the case of the compressor speed.

その後、コントローラ１０は、高圧圧力ＨＰの制御偏差ΔＨＰの絶対値｜ΔＨＰ｜が所定値ε３未満であるか否かを判断する（ステップＳ１０）。絶対値｜ΔＨＰ｜が所定値ε３未満でない場合（ステップＳ１０，Ｎｏ）には、室外ファン３の回転数が過小または過大であると判断し、室外ファン３の回転数を変更し(ステップＳ１１）、本処理を終了する。一方、絶対値｜ΔＨＰ｜が所定値ε３未満である場合（ステップＳ１０，Ｙｅｓ）には、そのまま本処理を終了する。   Thereafter, the controller 10 determines whether or not the absolute value | ΔHP | of the control deviation ΔHP of the high pressure HP is less than a predetermined value ε3 (step S10). If the absolute value | ΔHP | is not less than the predetermined value ε3 (step S10, No), it is determined that the rotational speed of the outdoor fan 3 is too small or excessive, and the rotational speed of the outdoor fan 3 is changed (step S11). This process is terminated. On the other hand, when the absolute value | ΔHP | is less than the predetermined value ε3 (step S10, Yes), this process is ended as it is.

ここで、所定値ε３は、制御を実施する最低偏差を示しており、図６を用いて、この最低偏差について説明する。図６は、凝縮器出口温度４５℃、蒸発温度７℃、蒸発器出口過熱度５℃、断熱効率０．７の条件における高圧圧力に対する理論ＣＯＰの変化を示す図である。図６により、高圧圧力が「Ａ」の範囲では、ＣＯＰの変化は１％以内であり、高圧圧力を「Ａ」の範囲、すなわち１１ＭＰａから１１．６ＭＰａまでの範囲に制御することによって、実用上、最適ＣＯＰと同等の性能を得ることができる。従って、所定値ε３としては、高圧圧力範囲「Ａ」の１／２、すなわち０．３ＭＰａ程度とすればよいことがわかる。   Here, the predetermined value ε3 indicates the minimum deviation for performing the control, and this minimum deviation will be described with reference to FIG. FIG. 6 is a graph showing changes in the theoretical COP with respect to high pressure under conditions of a condenser outlet temperature of 45 ° C., an evaporation temperature of 7 ° C., an evaporator outlet superheat degree of 5 ° C., and an adiabatic efficiency of 0.7. According to FIG. 6, when the high pressure is in the range of “A”, the change in COP is within 1%. By controlling the high pressure to be in the range of “A”, that is, from 11 MPa to 11.6 MPa, it is practically possible. The performance equivalent to the optimum COP can be obtained. Therefore, it is understood that the predetermined value ε3 may be set to ½ of the high pressure range “A”, that is, about 0.3 MPa.

室外ファン回転数の変更方法は、たとえばΔＨＰ＞＋ε３の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ*に比して高いため、室外ファン３の回転数が過小であると判断し、室外ファン３の回転数を増加させる。一方、ΔＨＰ＜−ε３の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ*に比して低いため、室外ファン３の回転数が過大であると判断し、室外ファン３の回転数を減少させる。室外ファン回転数の増加や減少の割合は、圧縮機回転数の場合と同様に、制御偏差の絶対値｜ΔＨＰ｜に応じて変更するようにすればよい。なお、上述した一連の制御処理は繰り返し行われる。   For example, when ΔHP> + ε3, the outdoor fan rotational speed is changed by determining that the rotational speed of the outdoor fan 3 is too low because the high pressure HP is higher than the target high pressure HP *. Increase the number of revolutions of 3. On the other hand, when ΔHP <−ε3, since the high pressure HP is lower than the target high pressure HP *, it is determined that the rotational speed of the outdoor fan 3 is excessive, and the rotational speed of the outdoor fan 3 is decreased. . The rate of increase or decrease in the outdoor fan rotation speed may be changed according to the absolute value | ΔHP | of the control deviation, as in the case of the compressor rotation speed. The series of control processes described above are repeated.

つぎに、図７に示すフローチャートを参照して、コントローラ１０による暖房運転時の制御処理手順について説明する。この空気調和機における暖房運転の制御処理は、暖房能力の確保に必要な高圧圧力を圧縮機１の回転数によって制御し、放熱器出口状態を電子式膨張弁４の開度によって制御し、さらに暖房能力の確保に必要な高圧圧力が最適高圧圧力に比して低い場合、高圧圧力を室外ファン７の回転数によって制御するものである。上述した冷房運転時と同様に、コントローラ１０のメモリ１０ａには、予め、放熱器出口温度ＣＴ０と冷凍サイクルの成績係数ＣＯＰが最大付近となる目標高圧圧力ＨＰ*との特性式を設定しておく。   Next, with reference to a flowchart shown in FIG. 7, a control processing procedure during heating operation by the controller 10 will be described. The control processing of the heating operation in this air conditioner is performed by controlling the high pressure required for ensuring the heating capacity by the number of revolutions of the compressor 1, controlling the radiator outlet state by the opening of the electronic expansion valve 4, and When the high pressure required for ensuring the heating capacity is lower than the optimum high pressure, the high pressure is controlled by the rotational speed of the outdoor fan 7. Similarly to the cooling operation described above, the memory 10a of the controller 10 is set in advance with a characteristic equation between the radiator outlet temperature CT0 and the target high pressure HP * at which the coefficient of performance COP of the refrigeration cycle is near the maximum. .

コントローラ１０は、まず、配管温度検出器２１，２２から検出値ＣＴ０，Ｔｄを取得する（ステップＳ２１）。さらに、高圧圧力ＨＰと目標高圧圧力ＨＰ１*との偏差ΔＨＰ１を算出し（ステップＳ２２）、放熱器出口温度ＣＴ０と目標放熱器出口温度ＣＴ０*との偏差ΔＣＴ０を算出する（ステップＳ２３）。その後、メモリ１０ａに記憶された関係から、放熱器出口温度ＣＴ０に対応した目標高圧圧力ＨＰ*を算出する（ステップＳ２４）。その後、コントローラ１０は、配管温度検出器２２からの検出値を用いて演算した高圧圧力ＨＰと目標高圧圧力ＨＰ*との偏差ΔＨＰを求める（ステップＳ２５）。   First, the controller 10 acquires the detection values CT0 and Td from the pipe temperature detectors 21 and 22 (step S21). Further, a deviation ΔHP1 between the high pressure HP and the target high pressure HP1 * is calculated (step S22), and a deviation ΔCT0 between the radiator outlet temperature CT0 and the target radiator outlet temperature CT0 * is calculated (step S23). Thereafter, the target high pressure HP * corresponding to the radiator outlet temperature CT0 is calculated from the relationship stored in the memory 10a (step S24). Thereafter, the controller 10 obtains a deviation ΔHP between the high pressure HP calculated using the detection value from the pipe temperature detector 22 and the target high pressure HP * (step S25).

ここで、配管温度検出器２２の検出値を用いて高圧圧力ＨＰは冷房運転時と同様の処理によって求めることができ、目標高圧圧力ＨＰ１は、たとえば運転台数や外気温度の関数として設定される。さらに、放熱器出口温度ＣＴ０は、配管温度検出器２１からの検出値であり、目標放熱器出口温度ＣＴ０*は、たとえば必要暖房能力が確保されるように、たとえば３０〜４５℃の値が設定される。   Here, the high pressure HP can be obtained by the same processing as in the cooling operation using the detection value of the pipe temperature detector 22, and the target high pressure HP1 is set as a function of the number of operating units and the outside air temperature, for example. Further, the radiator outlet temperature CT0 is a detected value from the pipe temperature detector 21, and the target radiator outlet temperature CT0 * is set to a value of 30 to 45 ° C., for example, so as to ensure the necessary heating capacity. Is done.

その後、コントローラ１０は、高圧圧力ＨＰの制御偏差ΔＨＰ１の絶対値｜ΔＨＰ１｜が所定値ε４未満であるか否かを判断する（ステップＳ２６）。絶対値｜ΔＨＰ１｜が所定値ε４未満でない場合（ステップＳ２６，Ｎｏ）、暖房能力が不足または過剰であると判断し、圧縮機１の回転数を変更し(ステップＳ２７）、ステップＳ２８に移行する。一方、絶対値｜ΔＨＰ１｜が所定値ε４未満である場合（ステップＳ２６，Ｙｅｓ）には、そのままステップＳ２８に移行する。   Thereafter, the controller 10 determines whether or not the absolute value | ΔHP1 | of the control deviation ΔHP1 of the high pressure HP is less than a predetermined value ε4 (step S26). If the absolute value | ΔHP1 | is not less than the predetermined value ε4 (No in step S26), it is determined that the heating capacity is insufficient or excessive, the rotation speed of the compressor 1 is changed (step S27), and the process proceeds to step S28. . On the other hand, when the absolute value | ΔHP1 | is less than the predetermined value ε4 (Yes in step S26), the process proceeds to step S28 as it is.

ここで、所定値ε４は、制御を実施する最低偏差を示しており、たとえば０．０５ＭＰａである。圧縮機回転数の変更方法は、たとえばΔＨＰ１＞＋ε４の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ１*に比して高いため、暖房能力が過剰であると判断し、圧縮機１の回転数を減少させる。一方、ΔＨＰ１＜−ε４の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ１*に比して低いため、暖房能力が不足であると判断し、圧縮機１の回転数を増加させる。圧縮機１の回転数の増加や減少割合は、制御偏差の絶対値｜ΔＨＰ１｜の値に応じて変更するようにすればよい。   Here, the predetermined value ε4 indicates a minimum deviation for performing the control, and is, for example, 0.05 MPa. For example, in the case of ΔHP1> + ε4, the compressor rotation speed is changed by determining that the heating capacity is excessive because the high pressure HP is higher than the target high pressure HP1 *. Decrease. On the other hand, in the case of ΔHP1 <−ε4, since the high pressure HP is lower than the target high pressure HP1 *, it is determined that the heating capacity is insufficient, and the rotation speed of the compressor 1 is increased. The increase or decrease rate of the rotation speed of the compressor 1 may be changed according to the absolute value | ΔHP1 | of the control deviation.

その後、放熱器出口温度ＣＴ０の制御偏差ΔＣＴ０の絶対値｜ΔＣＴ０｜が所定値ε５未満であるか否かを判断する（ステップＳ２８）。絶対値｜ΔＣＴ０｜が所定値ε５未満でない場合（ステップＳ２８，Ｎｏ）、電子式膨張弁４の開度が過小または過大であると判断し、電子式膨張弁４の開度を変更し(ステップＳ２９)、ステップＳ３０に移行する。一方、絶対値｜ΔＣＴ０｜が所定値ε５未満である場合（ステップＳ２８，Ｙｅｓ）には、そのままステップＳ３０に移行する。   Thereafter, it is determined whether or not the absolute value | ΔCT0 | of the control deviation ΔCT0 of the radiator outlet temperature CT0 is less than a predetermined value ε5 (step S28). If the absolute value | ΔCT0 | is not less than the predetermined value ε5 (No in step S28), it is determined that the opening degree of the electronic expansion valve 4 is too small or too large, and the opening degree of the electronic expansion valve 4 is changed (step S29), the process proceeds to step S30. On the other hand, when the absolute value | ΔCT0 | is less than the predetermined value ε5 (Yes in step S28), the process proceeds to step S30 as it is.

ここで、所定値ε５は、制御を実施する最低偏差を示しており、たとえば０．５℃である。電子式膨張弁４の開度の変更方法は、たとえばΔＣＴ０＞＋ε５の場合には、放熱器出口温度ＣＴ０が目標放熱器出口温度ＣＴ０*に比して大きいため、電子式膨張弁４の開度が過大であると判断し、電子式膨張弁４の開度を減少させる。一方、ΔＣＴ０＜−ε５の場合には、放熱器出口温度ＣＴ０が目標放熱器出口温度ＣＴ０*に比して低いため、電子式膨張弁４の開度が過小であると判断し、電子式膨張弁４の開度を増加させる。電子式膨張弁開度の増加や減少の割合は、冷房運転時の場合と同様に制御偏差の絶対値｜ΔＣＴ０｜の値に応じて変更するようにすればよい。   Here, the predetermined value ε5 indicates the minimum deviation for performing the control, and is 0.5 ° C., for example. The method of changing the opening degree of the electronic expansion valve 4 is, for example, when ΔCT0> + ε5, the radiator outlet temperature CT0 is larger than the target radiator outlet temperature CT0 *. Is determined to be excessive, and the opening degree of the electronic expansion valve 4 is decreased. On the other hand, when ΔCT0 <−ε5, since the radiator outlet temperature CT0 is lower than the target radiator outlet temperature CT0 *, it is determined that the opening degree of the electronic expansion valve 4 is too small, and the electronic expansion is performed. The opening degree of the valve 4 is increased. The rate of increase or decrease of the electronic expansion valve opening may be changed according to the absolute value | ΔCT0 | of the control deviation as in the cooling operation.

その後、コントローラ１０は、必要な暖房能力を確保するための目標高圧圧力ＨＰ１*が成績係数ＣＯＰを最大付近とする高圧圧力範囲ＨＰ*以下であるか否かを判断する（ステップＳ３０）。目標高圧圧力ＨＰ１*が高圧圧力範囲ＨＰ*以下でない場合（ステップＳ３０，Ｎｏ）には、暖房能力確保を優先して最適高圧圧力ＨＰ*への制御は行わず、そのまま本処理を終了する。一方、目標高圧圧力ＨＰ１*が最適高圧圧力ＨＰ*以下である場合（ステップＳ３０，Ｙｅｓ）には、成績係数ＣＯＰを優先して高圧圧力ＨＰを最適高圧圧力ＨＰ*に制御する。   Thereafter, the controller 10 determines whether or not the target high pressure HP1 * for ensuring the necessary heating capacity is equal to or lower than the high pressure range HP * having the coefficient of performance COP close to the maximum (step S30). If the target high pressure HP1 * is not less than or equal to the high pressure range HP * (step S30, No), priority is given to ensuring the heating capacity, and the control to the optimum high pressure HP * is not performed, and this processing is terminated as it is. On the other hand, when the target high pressure HP1 * is equal to or lower than the optimum high pressure HP * (step S30, Yes), the high pressure HP is controlled to the optimum high pressure HP * giving priority to the coefficient of performance COP.

すなわち、コントローラ１０は、高圧圧力の制御偏差ΔＨＰの絶対値｜ΔＨＰ｜が所定値ε６未満であるか否かを判断し（ステップＳ３１）、絶対値｜ΔＨＰ｜が所定値ε６未満でない場合（ステップＳ３１，Ｎｏ）には、室外ファン３の回転数が過小または過大であると判断し、室外ファン３の回転数を変更し(ステップＳ３２)、本処理を終了し、絶対値｜ΔＨＰ｜が所定値ε６未満である場合（ステップＳ３１，Ｙｅｓ）には、そのまま本処理を終了する。   That is, the controller 10 determines whether or not the absolute value | ΔHP | of the control deviation ΔHP of the high pressure is less than the predetermined value ε6 (step S31), and if the absolute value | ΔHP | is not less than the predetermined value ε6 (step S31). In S31, No), it is determined that the rotational speed of the outdoor fan 3 is too small or too large, the rotational speed of the outdoor fan 3 is changed (step S32), this process is terminated, and the absolute value | ΔHP | If it is less than the value ε6 (step S31, Yes), this processing is terminated as it is.

ここで、所定値ε６は、制御を実施する最低偏差を示しており、冷房運転の制御の場合と同様に、たとえば０．３ＭＰａ程度である。この変更方法は、たとえばΔＨＰ＞＋ε６の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ*に比して高いため、室外ファン３の回転数が過小であると判断し、室外ファン３の回転数を増加させる。一方、ΔＨＰ＜−ε６の場合には、高圧圧力ＨＰが目標高圧圧力ＨＰ*に比して低いため、室外ファン３の回転数が過大であると判断し、室外ファン３の回転数を減少させる。室外ファン回転数の増加や減少の割合は、冷房運転時の場合と同様に、制御偏差の絶対値｜ΔＨＰ｜に応じて変更するようにすればよい。   Here, the predetermined value ε6 indicates the minimum deviation for performing the control, and is about 0.3 MPa, for example, as in the case of the control of the cooling operation. For example, in the case of ΔHP> + ε6, this change method determines that the rotational speed of the outdoor fan 3 is too low because the high pressure HP is higher than the target high pressure HP *, and the rotational speed of the outdoor fan 3 Increase. On the other hand, when ΔHP <−ε6, since the high pressure HP is lower than the target high pressure HP *, it is determined that the rotational speed of the outdoor fan 3 is excessive, and the rotational speed of the outdoor fan 3 is decreased. . The rate of increase or decrease in the outdoor fan rotation speed may be changed according to the absolute value | ΔHP | of the control deviation, as in the cooling operation.

このように上述した実施の形態１では、予め、高圧側熱交換器（冷房時は室外熱交換器２、暖房時は室内熱交換器６）の出口温度すなわち放熱器出口温度と、成績係数ＣＯＰが最大付近となる高圧圧力範囲ＨＰ*との特性を記憶しておき、この特性と、配管温度検出器２０〜２３で検出された各検出値とをもとに、冷房あるいは暖房能力を圧縮機１の回転数で制御し、熱交換器の状態を電子式膨張弁４の開度で制御し、高圧圧力を室外ファン３の回転数で制御するようにしている。なお、この実施の形態１では高圧圧力を室外ファン３の回転数によって制御するようにしているが、室内ファン７によって制御しても良く、あるいは室内ファン７と室外ファン３との双方を制御するようにしても良い。   As described above, in the first embodiment described above, the outlet temperature of the high-pressure side heat exchanger (the outdoor heat exchanger 2 during cooling, the indoor heat exchanger 6 during heating), that is, the outlet temperature of the radiator, and the coefficient of performance COP. Is stored in the vicinity of the maximum pressure range HP *, and on the basis of this characteristic and the detected values detected by the pipe temperature detectors 20 to 23, the cooling or heating capacity is determined by the compressor. The heat exchanger state is controlled by the opening degree of the electronic expansion valve 4, and the high pressure is controlled by the rotational speed of the outdoor fan 3. In the first embodiment, the high pressure is controlled by the rotational speed of the outdoor fan 3, but it may be controlled by the indoor fan 7, or both the indoor fan 7 and the outdoor fan 3 are controlled. You may do it.

また、冷房および暖房能力を電子式膨張弁４の開度で制御し、熱交換器の状態を室内ファン７あるいは室外ファン３のうち少なくともどちらか一方の回転数で制御し、高圧圧力を圧縮機１の回転数で制御するようにしても良い。   In addition, the cooling and heating capacity is controlled by the opening degree of the electronic expansion valve 4, the state of the heat exchanger is controlled by the rotational speed of at least one of the indoor fan 7 and the outdoor fan 3, and the high pressure is compressed by the compressor. You may make it control by the rotation speed of 1.

さらに、冷房および暖房能力を室内ファン７あるいは室外ファン３のうち少なくともどちらか一方の回転数で制御し、熱交換器の状態を圧縮機１の回転数で制御し、高圧圧力を電子式膨張弁４の開度で制御するようにしても良い。   Further, the cooling and heating capacity is controlled by the rotational speed of at least one of the indoor fan 7 or the outdoor fan 3, the state of the heat exchanger is controlled by the rotational speed of the compressor 1, and the high pressure is controlled by an electronic expansion valve. You may make it control by the opening degree of 4.

また、冷房あるいは暖房能力を確保する運転を優先するか、成績係数ＣＯＰの高い運転を優先するかによって、上述した制御処理を切り替えるようにしてもよい。   Further, the above-described control processing may be switched depending on whether priority is given to an operation that ensures cooling or heating capability or an operation that has a high coefficient of performance COP.

この実施の形態１によれば、放熱器出口温度の値に応じて高圧圧力を、成績係数ＣＯＰが最大付近となる高圧圧力範囲に制御するようにしたため、常に成績係数ＣＯＰの高い運転を行うことができ、省エネルギー化に貢献できる超臨界冷凍サイクルを用いた空気調和機を実現することができる。   According to the first embodiment, the high pressure is controlled in the high pressure range in which the coefficient of performance COP is close to the maximum according to the value of the radiator outlet temperature. Therefore, the operation with a high coefficient of performance COP is always performed. Therefore, an air conditioner using a supercritical refrigeration cycle that can contribute to energy saving can be realized.

この場合、図１５に示した従来の空気調和機のように、被加熱（あるいは被冷却）流体である空気の流入温度に基づいた制御を行わないため、室外ファン３あるいは室内ファン７の風量が変化した場合であっても、高圧圧力範囲で運転することができる超臨界冷凍サイクルを用いた空気調和機を実現することができる。   In this case, unlike the conventional air conditioner shown in FIG. 15, since control based on the inflow temperature of air that is a heated (or cooled) fluid is not performed, the air volume of the outdoor fan 3 or the indoor fan 7 is Even if it changes, the air conditioner using the supercritical refrigeration cycle which can be operated in the high pressure range can be realized.

また、圧縮機吐出部から高圧側熱交換器に至る配管温度の検出値のみで高圧圧力を推定するようにしたため、圧力を直接測定する高価な圧力検出器などを用いる必要がなく、簡易な構成で安価な超臨界冷凍サイクルを用いた空気調和機を実現することができる。   In addition, since the high pressure is estimated only from the detected value of the piping temperature from the compressor discharge section to the high pressure side heat exchanger, there is no need to use an expensive pressure detector that directly measures the pressure, and a simple configuration And an inexpensive air conditioner using a supercritical refrigeration cycle can be realized.

実施の形態２．
つぎに、この発明の実施の形態２について説明する。図８は、この実施の形態２である冷凍サイクル装置の構成を示す図である。図８において、この冷凍サイクル装置は、空気調和機であり、実施の形態１と同様に、冷媒として、高圧側が臨界圧力以上となる二酸化炭素のような超臨界冷媒を用いている。この空気調和機は、図１５に示した空気調和機と同様に、室内熱交換器６、室内ファン７を含む室内ユニットＡと、室外熱交換器２、室外ファン３、圧縮機１およびアキュムレータ１２を含む室外ユニットＢとを有し、室内ユニットＡと室外ユニットＢとは、接続配管（液配管）５および接続配管（ガス配管）８によって接続されている。室内熱交換器６および室外熱交換器２を接続する他方の液配管５には、第２の冷媒回路であって放熱器（冷房時は室外熱交換器２、暖房時は室内熱交換器６）出口の冷媒温度を更に低下させる第２の放熱器３３と、絞り手段としての電子式膨張弁４とを含むブリッジ回路Ｃが設けられている。なお、図８に示した空気調和機では、室内ユニットＡが１台だけ接続された場合を示しているが、１台の室外ユニットＢに対して、室内ユニットＡが複数台設置される構成としても良い。 Embodiment 2. FIG.
Next, a second embodiment of the present invention will be described. FIG. 8 is a diagram showing the configuration of the refrigeration cycle apparatus according to the second embodiment. In FIG. 8, this refrigeration cycle apparatus is an air conditioner, and as in the first embodiment, a supercritical refrigerant such as carbon dioxide whose high pressure side is equal to or higher than the critical pressure is used as the refrigerant. This air conditioner includes an indoor unit A including an indoor heat exchanger 6 and an indoor fan 7, an outdoor heat exchanger 2, an outdoor fan 3, a compressor 1, and an accumulator 12, similarly to the air conditioner shown in FIG. 15. The indoor unit A and the outdoor unit B are connected by a connection pipe (liquid pipe) 5 and a connection pipe (gas pipe) 8. The other liquid pipe 5 that connects the indoor heat exchanger 6 and the outdoor heat exchanger 2 is a second refrigerant circuit that is a radiator (the outdoor heat exchanger 2 during cooling and the indoor heat exchanger 6 during heating. ) A bridge circuit C including a second radiator 33 that further lowers the refrigerant temperature at the outlet and an electronic expansion valve 4 as a throttle means is provided. In addition, although the case where only one indoor unit A is connected is shown in the air conditioner shown in FIG. 8, as a configuration in which a plurality of indoor units A are installed for one outdoor unit B Also good.

ブリッジ回路Ｃ内の第２の放熱器３３は、第２の圧縮機３０、第２の凝縮器３１、第２の絞り手段３２、およびこれらを接続する配管から構成される第２の冷媒回路に接続される。第２の放熱器３３内の第２の冷媒回路は、主冷媒回路の放熱器出口冷媒から熱を奪い、放熱器出口冷媒を冷却するとともに、第２の冷媒回路の冷媒自体は蒸発してガス冷媒となる蒸発器として機能する。   The second radiator 33 in the bridge circuit C is a second refrigerant circuit composed of the second compressor 30, the second condenser 31, the second throttling means 32, and piping connecting them. Connected. The second refrigerant circuit in the second radiator 33 takes heat from the radiator outlet refrigerant of the main refrigerant circuit, cools the radiator outlet refrigerant, and evaporates the refrigerant itself of the second refrigerant circuit. Functions as an evaporator serving as a refrigerant.

なお、この実施の形態２では、室外ユニットＢの小型化を図るため、凝縮器２と第２の凝縮器３１とは、室外ファン３を共用する構成としている。ここで、第２の冷媒回路の必要冷却能力は、主冷媒回路の冷房能力に比べて十分小さく、たとえば１０ＨＰ程度の空気調和機に対して１〜２馬力程度の空気調和機、すなわち主冷媒回路に比べて成績係数ＣＯＰの高い冷凍サイクル装置が使用される。   In the second embodiment, in order to reduce the size of the outdoor unit B, the condenser 2 and the second condenser 31 share the outdoor fan 3. Here, the required cooling capacity of the second refrigerant circuit is sufficiently smaller than the cooling capacity of the main refrigerant circuit, for example, an air conditioner of about 1 to 2 horsepower with respect to an air conditioner of about 10 HP, that is, the main refrigerant circuit. A refrigeration cycle apparatus having a higher coefficient of performance COP is used.

さらに、ブリッジ回路Ｃには、第２の放熱器３３から電子式膨張弁４への冷媒の流れが一方向となるように４つの逆止弁２５〜２８が設けられている。第２の放熱器３３としては、たとえば二重管式熱交換器やプレート熱交換器などが用いられ、第２の放熱器３３内では、放熱器出口の冷媒と第２の冷媒回路を流れる冷媒とが対向流となるように冷媒配管が接続される。   Further, the bridge circuit C is provided with four check valves 25 to 28 so that the refrigerant flows from the second radiator 33 to the electronic expansion valve 4 in one direction. As the second radiator 33, for example, a double tube heat exchanger, a plate heat exchanger, or the like is used. In the second radiator 33, the refrigerant flowing through the refrigerant at the outlet of the radiator and the second refrigerant circuit. Are connected to the refrigerant pipe so that they are opposed to each other.

ここで、この冷凍サイクル装置の動作について説明すると、まず、冷房運転時において、四方切替弁９は、図８の実線で示すように接続され、圧縮機１と室外熱交換器２とが連通し、室内熱交換器６および室外熱交換器２はそれぞれ蒸発器および放熱器として機能する。すなわち、圧縮機１から吐出された高温・高圧の冷媒ガスが室外熱交換器２に導入され、ここで、冷媒ガスと室外空気との熱交換が行われた後、中温・高圧ガスは逆止弁２７を通過して第２の放熱器３３に流入する。なお、このとき、逆止弁２８は高低圧力差で閉止される。   Here, the operation of the refrigeration cycle apparatus will be described. First, during the cooling operation, the four-way switching valve 9 is connected as shown by the solid line in FIG. 8, and the compressor 1 and the outdoor heat exchanger 2 communicate with each other. The indoor heat exchanger 6 and the outdoor heat exchanger 2 function as an evaporator and a radiator, respectively. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the outdoor heat exchanger 2, where after the heat exchange between the refrigerant gas and the outdoor air is performed, the medium-temperature and high-pressure gas is non-returned. It passes through the valve 27 and flows into the second radiator 33. At this time, the check valve 28 is closed with a high and low pressure difference.

第２の放熱器３３に流入した冷媒は、第２の冷媒回路によって熱を奪われ、冷媒自体の温度が低下した後、電子式膨張弁４によって減圧されて低温・低圧の二相冷媒となって、逆止弁２６を通過する。なお、このとき、逆止弁２５は高低圧力差で閉止される。ブリッジ回路Ｃを出た冷媒は、液配管５を通って室内熱交換器６に導入され、室内空気と熱交換が行われた後、再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→室外熱交換器２→逆止弁２７→第２の放熱器３３→電子式膨張弁４→逆止弁２６→液配管５→室内熱交換器６→ガス配管８→四方切替弁９→アキュムレータ１２→圧縮機１の順で主冷媒回路内を流動し、冷房運転が行われる。   The refrigerant flowing into the second radiator 33 is deprived of heat by the second refrigerant circuit, and after the temperature of the refrigerant itself is lowered, the refrigerant is decompressed by the electronic expansion valve 4 to become a low-temperature / low-pressure two-phase refrigerant. Pass through the check valve 26. At this time, the check valve 25 is closed with a high / low pressure difference. The refrigerant that has exited the bridge circuit C is introduced into the indoor heat exchanger 6 through the liquid pipe 5, exchanges heat with the indoor air, and then is sucked into the compressor 1 again. That is, the refrigerant is the compressor 1, the four-way switching valve 9, the outdoor heat exchanger 2, the check valve 27, the second radiator 33, the electronic expansion valve 4, the check valve 26, the liquid pipe 5, and the indoor heat exchange. The air flows through the main refrigerant circuit in the order of the vessel 6 → the gas pipe 8 → the four-way switching valve 9 → the accumulator 12 → the compressor 1, and the cooling operation is performed.

一方、暖房運転時において、四方切替弁９は、図８の点線で示すように接続され、圧縮機１と室内熱交換器６とが連通し、室内熱交換器６および室外熱交換器２はそれぞれ放熱器および蒸発器として機能する。すなわち、圧縮機１から吐出された高温・高圧の冷媒ガスが室内熱交換器６に導入され、ここで、冷媒ガスと室外空気との熱交換が行われた後、中温・高圧ガスは液配管５を通過した後、逆止弁２５を通過して第２の放熱器３３に流入する。なお、このとき、逆止弁２６は高低圧力差で閉止される。   On the other hand, during the heating operation, the four-way switching valve 9 is connected as shown by the dotted line in FIG. 8, the compressor 1 and the indoor heat exchanger 6 communicate with each other, and the indoor heat exchanger 6 and the outdoor heat exchanger 2 are Each functions as a radiator and an evaporator. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the indoor heat exchanger 6 where heat exchange between the refrigerant gas and the outdoor air is performed, and then the intermediate-temperature and high-pressure gas is liquid piping. After passing 5, it passes through the check valve 25 and flows into the second radiator 33. At this time, the check valve 26 is closed with a high and low pressure difference.

第２の放熱器３３に流入した冷媒は、第２の冷媒回路によって熱を奪われ、冷媒自体の温度を低下した後、電子式膨張弁４によって減圧されて低温・低圧の二相冷媒となって、逆止弁２８を通過する。なお、このとき、逆止弁２７は高低圧力差で閉止される。ブリッジ回路Ｃを出た冷媒は、室外熱交換器２に導入されて室外空気と熱交換が行われ、四方切替弁９およびアキュムレータ１２を通過して再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→室内熱交換器６→液配管５→逆止弁２５→第２の放熱器３３→電子式膨張弁４→逆止弁２８→室外熱交換器２→四方切替弁９→アキュムレータ１２→圧縮機１の順で主冷媒回路内を流動し、暖房運転が行われる。   The refrigerant flowing into the second radiator 33 is deprived of heat by the second refrigerant circuit, and after the temperature of the refrigerant itself is lowered, the refrigerant is decompressed by the electronic expansion valve 4 to become a low-temperature / low-pressure two-phase refrigerant. Pass through the check valve 28. At this time, the check valve 27 is closed with a high / low pressure difference. The refrigerant exiting the bridge circuit C is introduced into the outdoor heat exchanger 2 to exchange heat with outdoor air, passes through the four-way switching valve 9 and the accumulator 12, and is sucked into the compressor 1 again. That is, the refrigerant is the compressor 1, the four-way switching valve 9, the indoor heat exchanger 6, the liquid pipe 5, the check valve 25, the second radiator 33, the electronic expansion valve 4, the check valve 28, and the outdoor heat exchange. Heating operation is performed by flowing in the main refrigerant circuit in the order of the unit 2 → the four-way switching valve 9 → the accumulator 12 → the compressor 1.

この実施の形態２では、放熱器を出た超臨界流体を第２の放熱器３３によって更に冷却するようにしているため、成績係数ＣＯＰの高い運転が可能となる。この成績係数ＣＯＰが高くなる理由について、図９を参照して説明する。図９は、横軸に放熱器(フロン系冷媒Ｒ２２の場合は凝縮器)出口温度［℃］をとり、縦軸に成績係数ＣＯＰを示している。図９の実線は、冷房の標準的な条件で放熱器(Ｒ２２の場合は凝縮器)出口温度に対する成績係数ＣＯＰの変化を示し、Ｒ２２の場合とＣＯ2の場合とで比較した計算結果である。図９に示すように、Ｒ２２の凝縮器出口温度の低下量に対する成績係数ＣＯＰの増加率（傾き）は、ＣＯ2の傾きに対して小さく、放熱器（凝縮器）出口温度が２５℃以下の領域では、ＣＯ2の成績係数ＣＯＰがＲ２２を上回る結果となる。これは、Ｒ２２の場合、圧力−エンタルピー線図の特性上から凝縮器出口温度の低下量に対する蒸発器内エンタルピーの増加率がＣＯ2の場合に比べて小さいためである。 In the second embodiment, since the supercritical fluid exiting the radiator is further cooled by the second radiator 33, operation with a high coefficient of performance COP is possible. The reason why the coefficient of performance COP increases will be described with reference to FIG. In FIG. 9, the horizontal axis represents the outlet temperature [° C.] of the radiator (condenser in the case of the fluorocarbon refrigerant R22), and the vertical axis represents the coefficient of performance COP. The solid line in FIG. 9, (in the case of R22 condenser) radiator in standard conditions of cooling illustrates the variation of the COP for the outlet temperature is the calculation result of comparison between cases of R22 and CO 2 . As shown in FIG. 9, the increase rate (slope) of the coefficient of performance COP with respect to the decrease in the condenser outlet temperature of R22 is small with respect to the slope of CO 2 , and the radiator (condenser) outlet temperature is 25 ° C. or less. In the region, the CO 2 coefficient of performance COP exceeds R22. This is because, in the case of R22, the increase rate of the enthalpy in the evaporator with respect to the amount of decrease in the condenser outlet temperature is smaller than the case of CO 2 due to the characteristics of the pressure-enthalpy diagram.

放熱器出口の冷媒を更に冷却する手法として、従来から放熱器出口部と蒸発器出口部とを熱交換させる手法が知られているが、この手法には限界値が存在し、放熱器（凝縮器）出口温度を２５℃以下にすることができず、Ｒ２２の成績係数ＣＯＰを上回ることが難しいとされていた。しかし、この実施の形態２では、放熱器出口部を第２の冷媒回路によって冷却するため、放熱器（凝縮器）出口温度を２５℃以下にすることができ、成績係数ＣＯＰの高い超臨界冷凍サイクルを用いた冷凍サイクル装置を実現することができる。   As a method for further cooling the refrigerant at the outlet of the radiator, a method of exchanging heat between the outlet of the radiator and the outlet of the evaporator is conventionally known. However, this method has a limit value, and the radiator (condensation) The outlet temperature could not be 25 ° C. or lower, and it was difficult to exceed the coefficient of performance COP of R22. However, in the second embodiment, since the radiator outlet is cooled by the second refrigerant circuit, the radiator (condenser) outlet temperature can be 25 ° C. or lower, and the supercritical refrigeration with a high coefficient of performance COP is achieved. A refrigeration cycle apparatus using a cycle can be realized.

ここで、この実施の形態２では、成績係数ＣＯＰの高い超臨界冷凍サイクルを用いた冷凍サイクル装置を実現するために、第２の冷媒回路を用いているが、この第２の冷媒回路による第２の冷凍サイクルの成績係数ＣＯＰは、超臨界冷凍サイクルの成績係数ＣＯＰに比べて十分大きく、第２の冷媒回路の追加による消費電力の増加量に比べて超臨界冷凍サイクルの消費電力の低減量が大きいため、超臨界冷凍サイクルと第２の冷凍サイクルとを合わせたシステム全体としての成績係数ＣＯＰが高い冷凍サイクル装置を実現することができる。   Here, in the second embodiment, the second refrigerant circuit is used in order to realize a refrigeration cycle apparatus using a supercritical refrigeration cycle having a high coefficient of performance COP. The coefficient of performance COP of the refrigeration cycle 2 is sufficiently larger than the coefficient of performance COP of the supercritical refrigeration cycle, and the reduction in power consumption of the supercritical refrigeration cycle compared to the increase in power consumption due to the addition of the second refrigerant circuit. Therefore, it is possible to realize a refrigeration cycle apparatus having a high coefficient of performance COP as the entire system including the supercritical refrigeration cycle and the second refrigeration cycle.

この実施の形態２によれば、放熱器出口を出た超臨界流体を、第２の冷凍サイクルによって更に冷却するようにしているため、放熱器出口温度を大幅に低下させることができ、成績係数ＣＯＰの高い超臨界冷凍サイクルを有した冷凍サイクル装置を実現することができる。   According to the second embodiment, since the supercritical fluid that has exited the radiator outlet is further cooled by the second refrigeration cycle, the radiator outlet temperature can be greatly reduced, and the coefficient of performance A refrigeration cycle apparatus having a supercritical refrigeration cycle with a high COP can be realized.

実施の形態３．
つぎに、この発明の実施の形態３について説明する。上述した実施の形態２では、第２の冷凍サイクルを用いていたが、この実施の形態３では、この第２の冷凍サイクルの代替として低温の蓄熱媒体を用いるようにしている。 Embodiment 3 FIG.
Next, a third embodiment of the present invention will be described. In the second embodiment described above, the second refrigeration cycle is used. In the third embodiment, a low-temperature heat storage medium is used as an alternative to the second refrigeration cycle.

図１０は、この発明の実施の形態３である冷凍サイクル装置の構成を示す図である。図１０において、この冷凍サイクル装置は、空気調和機であり、実施の形態１と同様に、冷媒として、高圧側が臨界圧力以上となる二酸化炭素のような超臨界冷媒を用いている。図１０において、ブリッジ回路Ｃ内の第２の放熱器３３には、搬送ポンプ３４から蓄熱槽３５内に収められた蓄熱媒体３６が供給される。ここで、蓄熱槽３５内の蓄熱媒体３６は、冷却コイル３７によって冷却され、蓄熱槽３５としては、たとえば氷蓄熱槽が用いられ、蓄熱媒体３６としては、氷水、水、ブラインなどが用いられる。その他の構成は、実施の形態２と同じであり、同一構成部分には同一符号を付している。   FIG. 10 is a diagram showing a configuration of a refrigeration cycle apparatus according to Embodiment 3 of the present invention. In FIG. 10, this refrigeration cycle apparatus is an air conditioner, and as in the first embodiment, a supercritical refrigerant such as carbon dioxide whose high pressure side is equal to or higher than the critical pressure is used as the refrigerant. In FIG. 10, the heat storage medium 36 stored in the heat storage tank 35 is supplied from the transport pump 34 to the second radiator 33 in the bridge circuit C. Here, the heat storage medium 36 in the heat storage tank 35 is cooled by the cooling coil 37, and for example, an ice heat storage tank is used as the heat storage tank 35, and ice water, water, brine, or the like is used as the heat storage medium 36. Other configurations are the same as those of the second embodiment, and the same reference numerals are given to the same components.

搬送ポンプ３４から第２の放熱器３３に送り込まれた蓄熱媒体３６は、主冷媒回路の放熱器出口冷媒から熱を奪い、放熱器出口冷媒を冷却するとともに、蓄熱媒体自体は加熱されて温度が上昇し、蓄熱槽３５内に返送される。ここで、蓄熱槽３５内の蓄熱媒体３６は、夜間電力などを利用して冷却コイル３７によって冷却されて冷熱が蓄えられ、昼間の冷房などに利用されるものである。   The heat storage medium 36 sent from the transfer pump 34 to the second radiator 33 takes heat from the radiator outlet refrigerant of the main refrigerant circuit, cools the radiator outlet refrigerant, and the heat storage medium itself is heated to increase the temperature. It rises and is returned to the heat storage tank 35. Here, the heat storage medium 36 in the heat storage tank 35 is cooled by the cooling coil 37 using nighttime power or the like to store cold heat, and is used for daytime cooling or the like.

この実施の形態３では、昼間の冷房負荷が小さい場合など、余剰冷熱が蓄積されている蓄熱槽３５内の蓄熱媒体３６を有効に利用する。すなわち、異なる系統間の熱負荷の不均等性を解消してシステム全体としての省エネルギー化に貢献するものである。   In the third embodiment, the heat storage medium 36 in the heat storage tank 35 in which excess cooling heat is accumulated is used effectively, such as when the daytime cooling load is small. That is, it contributes to energy saving as a whole system by eliminating the unevenness of heat load between different systems.

この実施の形態３によれば、放熱器を出た超臨界流体を、第２の放熱器３３によって更に冷却するため、実施の形態２と同様に、成績係数ＣＯＰの高い超臨界冷凍サイクルを用いた冷凍サイクル装置を実現することができる。また、第２の放熱器３３を冷却する冷熱源として、夜間電力などを利用して蓄熱された余剰冷熱を用いるため、蓄熱槽３５を含めたシステムとしてのエネルギーの有効利用に貢献することができる。   According to the third embodiment, since the supercritical fluid exiting the radiator is further cooled by the second radiator 33, a supercritical refrigeration cycle having a high coefficient of performance COP is used as in the second embodiment. The refrigeration cycle apparatus can be realized. Moreover, since the excessive cold heat stored using night electricity etc. is used as a cold heat source which cools the 2nd heat radiator 33, it can contribute to the effective utilization of the energy as a system including the heat storage tank 35. FIG. .

実施の形態４．
つぎに、この発明の実施の形態４について説明する。図１１は、この発明の実施の形態４である冷凍サイクル装置の構成を示す図である。図１１において、この冷凍サイクル装置は、空気調和機であり、実施の形態１と同様に、冷媒として、高圧側が臨界圧力以上となる二酸化炭素のような超臨界冷媒を用いている。 Embodiment 4 FIG.
Next, a fourth embodiment of the present invention will be described. FIG. 11 is a diagram showing a configuration of a refrigeration cycle apparatus according to Embodiment 4 of the present invention. In FIG. 11, this refrigeration cycle apparatus is an air conditioner, and as in the first embodiment, a supercritical refrigerant such as carbon dioxide whose high pressure side is equal to or higher than the critical pressure is used as the refrigerant.

図１１において、この空気調和機は、実施の形態１と同様に、室内熱交換器６や室内ファン７などが収納された室内ユニットＡと、室外熱交換器２、室外ファン３、圧縮機１およびアキュムレータ１２などが収納された室外ユニットＢとを有し、これらは接続配管（液配管）５および接続配管（ガス配管）８によって接続されている。圧縮機１は、四方切替弁９を介して、室内熱交換器６および室外熱交換器２に接続されるとともに、シャフト４５によって膨張機構３８と同軸で連結されている。また、室内熱交換器６および室外熱交換器２を接続する他方の液配管５には、膨張機構３８が設けられ、この膨張機構３８はシャフト４５によって圧縮機１と同軸で連結されている。その他の構成は実施の形態１と同じであり、同一構成部分には同一符号を付している。   In FIG. 11, this air conditioner includes an indoor unit A in which an indoor heat exchanger 6, an indoor fan 7, and the like are stored, an outdoor heat exchanger 2, an outdoor fan 3, and a compressor 1 as in the first embodiment. And an outdoor unit B in which an accumulator 12 and the like are housed, and these are connected by a connection pipe (liquid pipe) 5 and a connection pipe (gas pipe) 8. The compressor 1 is connected to the indoor heat exchanger 6 and the outdoor heat exchanger 2 via the four-way switching valve 9 and is coaxially connected to the expansion mechanism 38 by a shaft 45. The other liquid pipe 5 connecting the indoor heat exchanger 6 and the outdoor heat exchanger 2 is provided with an expansion mechanism 38, and the expansion mechanism 38 is coaxially connected to the compressor 1 by a shaft 45. Other configurations are the same as those of the first embodiment, and the same reference numerals are given to the same components.

図１２は、圧縮機１と膨張機構３８との連結構造を示す模式図である。図１２において、圧縮機１および膨張機構３８は、たとえば共にスクロール式の圧縮・膨張機構で構成されている。圧縮機１は、圧縮機ガス吸入口４２を外側に、圧縮機ガス吐出口４１を中央にそれぞれ有するものであり、旋回スクロール４４を図１２の矢印方向（右回り）に回転させ、圧縮機ガス吸入口４２から吸入したＣＯ2冷媒を固定スクロール４３との隙間容積を減少させて圧縮し、圧縮機ガス吐出口４１から吐出する構成となっている。 FIG. 12 is a schematic diagram illustrating a connection structure between the compressor 1 and the expansion mechanism 38. In FIG. 12, the compressor 1 and the expansion mechanism 38 are both constituted by, for example, a scroll type compression / expansion mechanism. The compressor 1 has a compressor gas suction port 42 on the outside and a compressor gas discharge port 41 on the center, and the orbiting scroll 44 is rotated in the direction of the arrow (clockwise) in FIG. The CO 2 refrigerant sucked from the suction port 42 is compressed by reducing the volume of the gap with the fixed scroll 43 and discharged from the compressor gas discharge port 41.

一方、膨張機構３８は、圧縮機１とは逆の構成、すなわち、膨張機構ガス吐出口５２を外側に、膨張機構ガス吸入口５１を内側にそれぞれ有し、旋回スクロール５４を図１２の矢印方向（左回り）に回転させ、膨張機構ガス吸入口５１から吸入したＣＯ2冷媒を固定スクロール５３との隙間容積を増加して膨張させ、膨張機構ガス吐出口５２から吐出する構成となっている。 On the other hand, the expansion mechanism 38 has a configuration opposite to that of the compressor 1, that is, the expansion mechanism gas discharge port 52 on the outside, the expansion mechanism gas suction port 51 on the inside, and the orbiting scroll 54 in the direction of the arrow in FIG. It is configured to rotate (counterclockwise), expand the CO 2 refrigerant sucked from the expansion mechanism gas suction port 51 by increasing the gap volume with the fixed scroll 53, and discharge it from the expansion mechanism gas discharge port 52.

ここで、この冷凍サイクル装置の動作について説明すると、まず、冷房運転時において、四方切替弁９は、図１１の実線に示すように接続され、圧縮機１と室外熱交換器２とが連通し、室内熱交換器６および室外熱交換器２はそれぞれ蒸発器および放熱器として機能する。すなわち、圧縮機１から吐出された高温・高圧の冷媒ガスが室外熱交換器２に導入され、ここで、冷媒ガスと室外空気との熱交換が行われた後、中温・高圧ガスは膨張機構３８によって減圧されて低温・低圧ガスとなる。この低温・低圧ガスは、液配管５を通って室内熱交換器６に導入され、室内空気と熱交換が行われた後、再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→室外熱交換器２→膨張機構３８→液配管５→室内熱交換器６→ガス配管８→四方切替弁９→アキュムレータ１２→圧縮機１の順で冷媒回路内を流動し、冷房運転が行われる。   Here, the operation of the refrigeration cycle apparatus will be described. First, during the cooling operation, the four-way switching valve 9 is connected as shown by the solid line in FIG. 11 so that the compressor 1 and the outdoor heat exchanger 2 communicate with each other. The indoor heat exchanger 6 and the outdoor heat exchanger 2 function as an evaporator and a radiator, respectively. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the outdoor heat exchanger 2, where after the heat exchange between the refrigerant gas and the outdoor air is performed, the medium-temperature and high-pressure gas is expanded into an expansion mechanism. The pressure is reduced by 38 to become a low temperature / low pressure gas. The low-temperature and low-pressure gas is introduced into the indoor heat exchanger 6 through the liquid pipe 5, exchanges heat with the room air, and then is sucked into the compressor 1 again. That is, the refrigerant is compressor 1 → four-way switching valve 9 → outdoor heat exchanger 2 → expansion mechanism 38 → liquid piping 5 → indoor heat exchanger 6 → gas piping 8 → four-way switching valve 9 → accumulator 12 → compressor 1 The refrigerant flows in order and the cooling operation is performed.

一方、暖房運転時において、四方切替弁９は、図１１の点線に示すように接続され、圧縮機１と室内熱交換器６とが連通し、室内熱交換器６および室外熱交換器２はそれぞれ放熱器および蒸発器として機能する。すなわち、圧縮機１から吐出された高温・高圧の冷媒ガスが室内熱交換器６に導入され、ここで、冷媒ガスと室内空気との熱交換が行われた後、中温・高圧ガスは液配管５を通過し、膨張機構３８によって減圧されて低温・低圧ガスとなる。この低温・低圧ガスは、室外熱交換器２に導入されて室外空気と熱交換が行われ、再び圧縮機１に吸入される。すなわち、冷媒は、圧縮機１→四方切替弁９→ガス配管８→室内熱交換器６→液配管５→膨張機構３８→室外熱交換器２→四方切替弁９→アキュムレータ１２→圧縮機１の順で冷媒回路内を流動し、暖房運転が行われる。   On the other hand, during the heating operation, the four-way switching valve 9 is connected as shown by a dotted line in FIG. 11 so that the compressor 1 and the indoor heat exchanger 6 communicate with each other, and the indoor heat exchanger 6 and the outdoor heat exchanger 2 are Each functions as a radiator and an evaporator. That is, the high-temperature and high-pressure refrigerant gas discharged from the compressor 1 is introduced into the indoor heat exchanger 6, where after the heat exchange between the refrigerant gas and the room air is performed, the medium-temperature and high-pressure gas is liquid pipe. 5 and is decompressed by the expansion mechanism 38 to become low-temperature and low-pressure gas. This low-temperature and low-pressure gas is introduced into the outdoor heat exchanger 2 to exchange heat with outdoor air, and is sucked into the compressor 1 again. That is, the refrigerant is the compressor 1 → the four-way switching valve 9 → the gas pipe 8 → the indoor heat exchanger 6 → the liquid pipe 5 → the expansion mechanism 38 → the outdoor heat exchanger 2 → the four-way switching valve 9 → the accumulator 12 → the compressor 1. In order, the refrigerant flows in the refrigerant circuit and the heating operation is performed.

この実施の形態４では、上述したように、圧縮機１の駆動軸と膨張機構３８の出力軸がシャフト４５で連結されており、膨張機構３８の出力を圧縮機１の駆動力の一部として利用する。すなわち、膨張機構３８の旋回スクロール３８の回転力を、膨張機構３８に連結する圧縮機１の駆動力の一部として利用することによって、圧縮機１の必要動力を低減することができ、成績係数ＣＯＰを向上させることができる。   In the fourth embodiment, as described above, the drive shaft of the compressor 1 and the output shaft of the expansion mechanism 38 are connected by the shaft 45, and the output of the expansion mechanism 38 is used as part of the drive force of the compressor 1. Use. That is, by using the rotational force of the orbiting scroll 38 of the expansion mechanism 38 as a part of the driving force of the compressor 1 connected to the expansion mechanism 38, the required power of the compressor 1 can be reduced, and the coefficient of performance COP can be improved.

また、冷媒としてＣＯ2を用いる場合、高圧圧力が超臨界圧力で運転されるため、膨張機構３８を用いることによって、蒸発器内のエンタルピー差を大きく増加させることができる。この蒸発器内のエンタルピー差の増加について、図１３を参照して説明する。 Further, when CO 2 is used as the refrigerant, the high pressure is operated at a supercritical pressure, and therefore the enthalpy difference in the evaporator can be greatly increased by using the expansion mechanism 38. The increase in the enthalpy difference in the evaporator will be described with reference to FIG.

図１３は、空調条件での標準的な運転動作を圧力−エンタルピー線図上に示したものであり、図１３（ａ）は、冷媒としてＲ２２を用いた場合の運転動作を示し、図１３（ｂ）は、冷媒としてＣＯ2を用いた場合の運転動作を示している。図１３（ａ）において、曲線Ｌ５は、絞り手段を用いた場合の減圧変化の様子を示し、曲線Ｌ６は、膨張機構を用いた場合の減圧変化の様子を示している。図１３（ａ）において、絞り手段を用いた場合は、ほぼ等エンタルピー変化となり、膨張機構を用いた場合は、ほぼ等エントロピー変化となる。絞り手段を膨張機構に変更することによって、蒸発器内のエンタルピー差は、ΔＨｈからΔＨｓに増加するが、Ｒ２２の場合にはこのエンタルピー差の増加率（＝（ΔＨｓ−ΔＨｈ）／ΔＨｈ×１００）が、高々１〜５％程度と小さいため、通常、冷凍サイクル中に膨張機構が導入されるケースはほとんど見られなかった。これは、高圧圧力と低圧圧力との圧力差（図１３（ａ）では、２０−６＝１４ｋｇ／ｃｍ2）が小さいためである。 FIG. 13 shows a standard operation operation under air-conditioning conditions on a pressure-enthalpy diagram. FIG. 13A shows an operation operation when R22 is used as a refrigerant. b) shows the operation when CO 2 is used as the refrigerant. In FIG. 13 (a), a curve L5 shows a change in pressure reduction when the throttle means is used, and a curve L6 shows a change in pressure reduction when the expansion mechanism is used. In FIG. 13A, when the diaphragm means is used, the change is substantially isenthalpy, and when the expansion mechanism is used, the change is substantially isentropic. By changing the throttling means to the expansion mechanism, the enthalpy difference in the evaporator increases from ΔHh to ΔHs. In the case of R22, the rate of increase of this enthalpy difference (= (ΔHs−ΔHh) / ΔHh × 100). However, since it is as small as about 1 to 5% at most, normally, the case where the expansion mechanism is introduced into the refrigeration cycle is hardly seen. This is because the pressure difference between the high pressure and the low pressure (20−6 = 14 kg / cm 2 in FIG. 13A) is small.

一方、図１３（ｂ）において、ＣＯ2の場合には、高圧圧力と低圧圧力との圧力差（図１３（ｂ）では、１１５−４２＝７３ｋｇ／ｃｍ2）が大きく、絞り手段を膨張機構に変更することによって、エンタルピー差の増加率が１０〜１５％程度大きくなる。したがって、膨張機構を導入することによって、Ｒ２２の場合に比べて成績係数ＣＯＰを大きく向上させることができ、省エネルギー化に大きく貢献することができる。 On the other hand, in FIG. 13B, in the case of CO 2 , the pressure difference between the high pressure and the low pressure (115−42 = 73 kg / cm 2 in FIG. 13B) is large, and the throttle means is an expansion mechanism. By changing to, the increase rate of the enthalpy difference increases by about 10 to 15%. Therefore, by introducing the expansion mechanism, the coefficient of performance COP can be greatly improved as compared with the case of R22, which can greatly contribute to energy saving.

なお、この実施の形態４では、膨張機構の出力を圧縮機の駆動力の一部として利用する例を示したが、圧縮機の駆動力ではなく、室外ファン３や室内ファン７の駆動力の一部として利用しても、成績係数ＣＯＰの高い超臨界冷凍サイクルをもった冷凍サイクル装置を実現することができる。   In the fourth embodiment, the output of the expansion mechanism is used as a part of the driving force of the compressor. However, the driving force of the outdoor fan 3 and the indoor fan 7 is not the driving force of the compressor. Even if it is used as a part, a refrigeration cycle apparatus having a supercritical refrigeration cycle with a high coefficient of performance COP can be realized.

この実施の形態４によれば、圧縮機１の駆動軸と膨張機構３８の出力軸とを連結させ、膨張機構３８の出力を圧縮機１の駆動力の一部として利用し、圧縮機動力を低減するとともに、蒸発器内のエンタルピー差を増加させことができるため、成績係数ＣＯＰの高い超臨界冷凍サイクルをもつ冷凍サイクル装置を実現することができる。   According to the fourth embodiment, the drive shaft of the compressor 1 and the output shaft of the expansion mechanism 38 are connected, the output of the expansion mechanism 38 is used as part of the drive force of the compressor 1, and the compressor power is used. While reducing the enthalpy difference in the evaporator, a refrigeration cycle apparatus having a supercritical refrigeration cycle with a high coefficient of performance COP can be realized.

実施の形態５．
つぎに、この発明の実施の形態５について説明する。図１４（ａ）〜（ｃ）は、横軸に冷媒質量速度Ｇ［ｋｇ／ｍ2ｓ］をとり、図１４（ａ）では縦軸に蒸発熱伝達率ｈｅ［Ｗ／ｍ2Ｋ］を示し、図１４（ｂ）では凝縮熱伝達率ｈｃ［Ｗ／ｍ2Ｋ］を示し、図１４（ｃ）では単位長さ当たりの圧力損失ΔＰ［（ｋｇ／ｃｍ2）／ｍ］を示したものである。図１４（ａ）〜（ｃ）は、標準的な空調用熱交換器の伝熱管内におけるＲ２２とＣＯ2との性能を比較した計算結果である。図１４において、「○」はＣＯ2の計算結果を示し、「●」はＲ２２の計算結果を示している。 Embodiment 5 FIG.
Next, a fifth embodiment of the present invention will be described. 14A to 14C, the horizontal axis represents the refrigerant mass velocity G [kg / m 2 s], and in FIG. 14A, the vertical axis represents the evaporation heat transfer coefficient he [W / m 2 K]. FIG. 14B shows the condensation heat transfer coefficient hc [W / m 2 K], and FIG. 14C shows the pressure loss ΔP [(kg / cm 2 ) / m] per unit length. Is. FIGS. 14A to 14C are calculation results comparing the performances of R22 and CO 2 in a heat transfer tube of a standard air conditioning heat exchanger. In FIG. 14, “◯” indicates the calculation result of CO 2 , and “●” indicates the calculation result of R 22.

図１４（ｃ）において、同一冷媒質量速度の条件下で、Ｒ２２とＣＯ2との圧力損失ΔＰを比較すると、ＣＯ2の圧力損失ΔＰは、Ｒ２２の圧力損失ΔＰに比べて大きく低下する。これは、ＣＯ2がフロン系冷媒に比べて高圧冷媒であり、蒸気密度がＲ２２の４〜５倍程度となるためである。換言すれば、ＣＯ2の圧力損失ΔＰをＲ２２の圧力損失ΔＰと同等にするには、たとえば冷媒質量速度Ｇを２５０ｋｇ／ｍ2ｓから４３０ｋｇ／ｍ2ｓまで、７０％程度増加させることができる（図１４（ｃ）の矢印Ｌ７参照）。このとき、図１４（ａ）に示すように、ＣＯ2の蒸発熱伝達率ｈｅは、Ｒ２２の蒸発伝達率ｈｅに比べて５０％程度増加し、図１４（ｂ）に示すように、凝縮熱伝達率ｈｃは、Ｒ２２の凝縮熱伝達率ｈｃとほぼ同等となるため、ＣＯ2はＲ２２に比べて高効率な熱交換器を実現することができる。 In FIG. 14C, when the pressure loss ΔP between R22 and CO 2 is compared under the condition of the same refrigerant mass velocity, the pressure loss ΔP of CO 2 greatly decreases compared to the pressure loss ΔP of R22. This is because CO 2 is a high-pressure refrigerant as compared with the fluorocarbon refrigerant, and the vapor density is about 4 to 5 times that of R22. In other words, in order to make the pressure loss ΔP of CO 2 equal to the pressure loss ΔP of R22, for example, the refrigerant mass velocity G can be increased by about 70% from 250 kg / m 2 s to 430 kg / m 2 s. (See arrow L7 in FIG. 14C). At this time, as shown in FIG. 14A, the evaporation heat transfer coefficient he of CO 2 is increased by about 50% compared to the evaporation transfer coefficient he of R22, and as shown in FIG. Since the transfer rate hc is substantially equal to the condensation heat transfer rate hc of R22, CO 2 can realize a heat exchanger that is more efficient than R22.

ところで、標準的な空調用熱交換器の伝熱管としては、管外径７ｍｍ程度の伝熱管が広く使用されている。したがって、実際の空気調和機では、多少の圧力損失増加を許容するとして、冷媒質量速度を７０％以上増加させるためには、熱交換器を構成する伝熱管の管外径を５ｍｍ以下にすれば良く、このような伝熱管を搭載することによって、ＣＯ2冷媒に適した熱交換器の小型化を図るようにしている。 By the way, as a heat transfer tube of a standard air conditioner heat exchanger, a heat transfer tube having an outer diameter of about 7 mm is widely used. Accordingly, in an actual air conditioner, assuming that a slight increase in pressure loss is allowed, in order to increase the refrigerant mass velocity by 70% or more, the outer diameter of the heat transfer tube constituting the heat exchanger should be 5 mm or less. Well, by mounting such a heat transfer tube, the heat exchanger suitable for the CO 2 refrigerant can be downsized.

また、上述したように、ＣＯ2は圧力損失の小さな冷媒であり、室内ユニットＡと室外ユニットＢとの接続配管であるガス配管８においても圧力損失が低下する。一方、液配管５においては、ＣＯ2の圧力損失はＲ２２の圧力損失に比べて増加する。これは、蒸気密度の場合とは逆に、ＣＯ2の液密度がＲ２２に比して低下するからである。たとえば、温度１８℃におけるＣＯ2の飽和液密度は、Ｒ２２に比して約３５％低下する。したがって、ＣＯ2を冷媒として用いる場合、ガス配管８の管径を液配管５の管径に比して大きくなるように構成するのではなく、ガス配管８と液配管５との管径を等しくなるように構成することによって、ＣＯ2に適した冷媒回路を構成することができる。また、ガス配管８と液配管５とを同一配管径とすることによって、ガス配管８と液配管５との区別が不必要となり、冷媒配管工事の作業性を向上させることができる。 Further, as described above, CO 2 is a refrigerant having a small pressure loss, and the pressure loss is also reduced in the gas pipe 8 that is a connecting pipe between the indoor unit A and the outdoor unit B. On the other hand, in the liquid pipe 5, the pressure loss of CO 2 increases compared to the pressure loss of R22. This is because, contrary to the case of the vapor density, the liquid density of CO 2 is lower than that of R22. For example, the saturated liquid density of CO 2 at a temperature of 18 ° C. is reduced by about 35% compared to R22. Therefore, when CO 2 is used as the refrigerant, the gas pipe 8 and the liquid pipe 5 are equal in diameter, rather than configured so that the pipe diameter of the gas pipe 8 is larger than the pipe diameter of the liquid pipe 5. By configuring so, a refrigerant circuit suitable for CO 2 can be configured. Further, by making the gas pipe 8 and the liquid pipe 5 have the same pipe diameter, the gas pipe 8 and the liquid pipe 5 need not be distinguished from each other, and the workability of the refrigerant pipe work can be improved.

ところで、フロン系冷媒では、たとえば大規模ビル用の空気調和機として利用される場合、室内ユニットＡと室外ユニットＢとの接続配管が長くなるため、冷房運転時にガス配管８での圧力損失が増大し、圧縮機１の吸入圧力が低下し、冷房能力や成績ＣＯＰなどの性能低下割合が大きくなる。特に、１５０ｍ以上の接続配管５，８を用いる場合には、それらの低下割合が１０％以上にも達し、大規模ビルなど、長い接続配管を必要とする空気調和機としては、性能低下量が許容範囲を超えるとされていた。しかしながら、ＣＯ2冷媒は、フロン系冷媒に比べてガス配管８での圧力損失を小さく抑えることができ、液配管５やガス配管８などの接続配管が１５０ｍ以上となる場合においても、性能低下が十分許容し得る大規模ビル対応の空気調和機を提供することができる。 By the way, in the case of using a CFC-based refrigerant, for example, as an air conditioner for a large-scale building, the connection pipe between the indoor unit A and the outdoor unit B becomes long, so that the pressure loss in the gas pipe 8 increases during the cooling operation. As a result, the suction pressure of the compressor 1 decreases, and the rate of performance reduction such as cooling capacity and performance COP increases. In particular, when connecting pipes 5 and 8 having a length of 150 m or more are used, the rate of the reduction reaches 10% or more. For an air conditioner that requires a long connecting pipe such as a large-scale building, the performance reduction amount is low. It was supposed to exceed the allowable range. However, the CO 2 refrigerant can suppress the pressure loss in the gas pipe 8 to be smaller than that of the chlorofluorocarbon refrigerant, and even when the connection pipe such as the liquid pipe 5 or the gas pipe 8 is 150 m or more, the performance deterioration is caused. An air conditioner for a large-scale building that is sufficiently acceptable can be provided.

この実施の形態５によれば、熱交換器を構成する伝熱管の管外径を５ｍｍ以下としたため、ＣＯ2冷媒の特性を利用して超臨界冷凍サイクルを用いた冷凍サイクル装置に適した熱交換器を構成することができるとともに、熱交換器の小型化を図ることができる。また、ガス配管８と液配管５との管径を等しくなるように構成したため、ＣＯ2に適した冷媒回路を構成することができるとともに、冷媒配管工事の作業性を向上させることができる。さらに、液配管５やガス配管８などの接続配管が１５０ｍ以上となる大規模ビルなどの空気調和機にＣＯ2を適用したため、性能低下が十分許容し得る大規模ビル対応の空気調和機を構成することができる。 According to the fifth embodiment, since the tube outer diameter of the heat transfer tube constituting the heat exchanger is set to 5 mm or less, the heat suitable for the refrigeration cycle apparatus using the supercritical refrigeration cycle utilizing the characteristics of the CO 2 refrigerant. While being able to constitute an exchanger, it is possible to reduce the size of the heat exchanger. In addition, since the gas pipe 8 and the liquid pipe 5 are configured to have the same diameter, a refrigerant circuit suitable for CO 2 can be formed, and workability of the refrigerant pipe work can be improved. Furthermore, because CO 2 is applied to air conditioners such as large-scale buildings where the connection pipes such as the liquid pipes 5 and the gas pipes 8 are 150 m or more, an air conditioner for large-scale buildings that can tolerate performance degradation is constructed. can do.

なお、上述した実施の形態１〜５では、空気である被加熱媒体あるいは被冷却媒体を制御する空気調和機である冷凍サイクル装置を示して説明したが、これに限らず、水あるいはブラインなどの被加熱媒体あるいは被冷却媒体を制御する冷凍サイクル装置にも適用することができる。   In the above-described first to fifth embodiments, the refrigeration cycle apparatus that is an air conditioner that controls a heated medium or a cooled medium that is air has been shown and described. The present invention can also be applied to a refrigeration cycle apparatus that controls a medium to be heated or a medium to be cooled.

この発明の実施の形態１である冷凍サイクル装置の冷媒回路図である。It is a refrigerant circuit diagram of the refrigeration cycle apparatus which is Embodiment 1 of this invention. 図１に示した冷媒回路の冷凍サイクルにおける圧力−エンタルピー線図である。It is a pressure-enthalpy diagram in the refrigerating cycle of the refrigerant circuit shown in FIG. 図１に示した冷媒回路の冷凍サイクルにおける高圧圧力と成績係数ＣＯＰとの関係を示す図である。It is a figure which shows the relationship between the high pressure in the refrigerating cycle of the refrigerant circuit shown in FIG. 1, and a coefficient of performance COP. 図１に示したコントローラによる冷房運転制御処理手順を示すフローチャートである。It is a flowchart which shows the cooling operation control processing procedure by the controller shown in FIG. 図１に示したコントローラによる高圧圧力の推定処理を説明する説明図である。It is explanatory drawing explaining the estimation process of the high pressure by the controller shown in FIG. 所定条件における高圧圧力に対する理論ＣＯＰの変化を示す図である。It is a figure which shows the change of theoretical COP with respect to the high pressure in a predetermined condition. 図１に示したコントローラによる暖房運転制御処理手順を示すフローチャートである。It is a flowchart which shows the heating operation control processing procedure by the controller shown in FIG. この発明の実施の形態２である冷凍サイクル装置の冷媒回路図である。It is a refrigerant circuit figure of the refrigerating cycle device which is Embodiment 2 of this invention. 図８に示した冷媒回路における凝縮器出口温度と成績係数ＣＯＰとの関係を示す図である。It is a figure which shows the relationship between the condenser exit temperature and the coefficient of performance COP in the refrigerant circuit shown in FIG. この発明の実施の形態３である冷凍サイクル装置の冷媒回路図である。It is a refrigerant circuit figure of the refrigerating cycle device which is Embodiment 3 of this invention. この発明の実施の形態４である冷凍サイクル装置の冷媒回路図である。It is a refrigerant circuit figure of the refrigerating cycle device which is Embodiment 4 of this invention. 図１１に示した圧縮機と膨張機構との連結状態を示す図である。It is a figure which shows the connection state of the compressor shown in FIG. 11, and an expansion mechanism. 圧力−エンタルピー線図上におけるＣＯ2冷媒とＲ２２冷媒との運転動作を比較した図である。Pressure - is a graph comparing the running operation of the CO 2 refrigerant and R22 refrigerant in diagram enthalpy. 冷媒質量速度に対する蒸発熱伝達率、凝縮熱伝達率、圧力損失の関係を、ＣＯ2冷媒とＲ２２冷媒とを比較して示す図である。Evaporation heat transfer coefficient for the coolant mass velocity, the condensation heat transfer coefficient, the relationship between the pressure loss, which is a graph showing by comparison the CO 2 refrigerant and R22 refrigerant. 従来の冷凍サイクル装置の冷媒回路図である。It is a refrigerant circuit diagram of the conventional refrigeration cycle apparatus.

符号の説明Explanation of symbols

１ 圧縮機、２ 室外熱交換器、３ 室外ファン、３ａ 室外ファンモータ、４ 絞り手段（電子式膨張弁）、５ 接続配管（液配管）、６ 室内熱交換器、７ 室内ファン、７ａ 室内ファンモータ、８ 接続配管（ガス配管）、９ 四方切替弁、１０ コントローラ、１０ａ メモリ、１１ 吐出圧力検出器、１２ アキュムレータ、１３ａ 流入管、１３ｂ 吸入配管、２０〜２３ 配管温度検出器、２５〜２８ 逆止弁、３０ 第２の圧縮機、３１ 第２の凝縮器、３２ 第２の絞り手段、３３ 第２の放熱器、３４ 搬送ポンプ、３５ 蓄熱槽、３６ 蓄熱媒体、３７ 冷却コイル、３８ 膨張機構、４１ 圧縮機ガス吐出口、４２ 圧縮機ガス吸入口、４３，５３ 固定スクロール、４４，５４ 旋回スクロール、４５ シャフト、５１ 膨張機構ガス吸入口、５２ 膨張機構ガス吐出口   DESCRIPTION OF SYMBOLS 1 Compressor, 2 Outdoor heat exchanger, 3 Outdoor fan, 3a Outdoor fan motor, 4 Throttle means (electronic expansion valve), 5 Connection piping (liquid piping), 6 Indoor heat exchanger, 7 Indoor fan, 7a Indoor fan Motor, 8 Connection piping (gas piping), 9 Four-way selector valve, 10 Controller, 10a Memory, 11 Discharge pressure detector, 12 Accumulator, 13a Inlet tube, 13b Suction piping, 20-23 Piping temperature detector, 25-28 Reverse Stop valve, 30 second compressor, 31 second condenser, 32 second throttle means, 33 second radiator, 34 transport pump, 35 heat storage tank, 36 heat storage medium, 37 cooling coil, 38 expansion mechanism , 41 Compressor gas discharge port, 42 Compressor gas suction port, 43, 53 Fixed scroll, 44, 54 Orbiting scroll, 45 Shaft, 51 Expansion mechanism gas Inlet, 52 expansion mechanism gas discharge port

Claims (6)

少なくとも圧縮機、高圧側熱交換器、膨張機構部、および低圧側熱交換器を液配管およびガス配管を用いて接続した冷媒回路を形成するとともに、前記高圧側熱交換器および前記低圧側熱交換器に対して被加熱媒体および被冷却媒体を搬送する被加熱媒体搬送手段および被冷却媒体搬送手段を有し、高圧側が超臨界圧力で運転される冷媒を用いた冷凍サイクル装置において、
前記膨張機構部の出力軸を、前記圧縮機、前記被加熱媒体搬送手段、前記被冷却媒体搬送手段のうち少なくとも１つの駆動軸に連結したことを特徴とする冷凍サイクル装置。 At least a compressor, a high pressure side heat exchanger, an expansion mechanism, and a low pressure side heat exchanger are formed using a liquid pipe and a gas pipe to form a refrigerant circuit, and the high pressure side heat exchanger and the low pressure side heat exchange are formed. In a refrigeration cycle apparatus using a refrigerant having a heated medium conveying means and a cooled medium conveying means for conveying a medium to be heated and a medium to be cooled with respect to a container, the high pressure side being operated at a supercritical pressure,
The refrigeration cycle apparatus characterized in that an output shaft of the expansion mechanism section is connected to at least one drive shaft among the compressor, the heated medium conveying means, and the cooled medium conveying means. 前記高圧側熱交換器の出口部を冷却する低温熱源をさらに備えたことを特徴とする請求項１に記載の冷凍サイクル装置。   The refrigeration cycle apparatus according to claim 1, further comprising a low-temperature heat source for cooling an outlet portion of the high-pressure side heat exchanger. 冷凍サイクルを行う第２冷凍サイクル手段をさらに備え、
前記低温熱源は、前記第２冷凍サイクル手段の蒸発器であることを特徴とする請求項２に記載の冷凍サイクル装置。 A second refrigeration cycle means for performing a refrigeration cycle;
The refrigeration cycle apparatus according to claim 2, wherein the low-temperature heat source is an evaporator of the second refrigeration cycle means. 前記高圧側熱交換器および前記低圧側熱交換器を形成する伝熱管の管外径を５ｍｍ以下とすることを特徴とする請求項１〜３のいずれか一つに記載の冷凍サイクル装置。   The refrigeration cycle apparatus according to any one of claims 1 to 3, wherein a tube outer diameter of a heat transfer tube forming the high-pressure side heat exchanger and the low-pressure side heat exchanger is 5 mm or less. 前記液配管および前記ガス配管を同一管径とすることを特徴とする請求項１〜４のいずれか一つに記載の冷凍サイクル装置。   The refrigeration cycle apparatus according to any one of claims 1 to 4, wherein the liquid pipe and the gas pipe have the same pipe diameter. 前記液配管および前記ガス配管の長さを１５０ｍ以上としたことを特徴とする請求項１〜５のいずれか一つに記載の冷凍サイクル装置。   The length of the said liquid piping and the said gas piping was 150 m or more, The refrigeration cycle apparatus as described in any one of Claims 1-5 characterized by the above-mentioned.
JP2005112540A 2005-04-08 2005-04-08 Refrigerating cycle device Pending JP2005249384A (en)

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