JP2006275399A - Air conditioner - Google Patents

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JP2006275399A
JP2006275399A JP2005094627A JP2005094627A JP2006275399A JP 2006275399 A JP2006275399 A JP 2006275399A JP 2005094627 A JP2005094627 A JP 2005094627A JP 2005094627 A JP2005094627 A JP 2005094627A JP 2006275399 A JP2006275399 A JP 2006275399A
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heat exchanger
air
load
indoor
refrigerant
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JP4617958B2 (en
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Makoto Saito
信 齊藤
Takashi Okazaki
多佳志 岡崎
Fumitake Unezaki
史武 畝崎
Tetsuji Nanatane
哲二 七種
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Mitsubishi Electric Corp
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Abstract

<P>PROBLEM TO BE SOLVED: To provide an air conditioning device applying natural refrigerant and significantly improving its efficiency. <P>SOLUTION: In this air conditioning device having a plurality of indoor heat exchangers for processing air conditioning load while dividing it into latent heat and sensible heat, air supply of a second heat exchanger is controlled to control heat treatment capacity of a first heat exchanger to be higher than that of the second heat exchanger, so that a temperature of the first heat exchanger at sensible heat side becomes higher than a dew point temperature of indoor air, and a temperature of the second heat exchanger becomes lower than the dew point temperature. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

この発明は、蒸気圧縮式冷凍サイクルによって冷媒を循環させ効率の良い空気調和を行う装置、例えば、室内温度と湿度を調整可能な空気調和機の運転効率向上に関するものである。   The present invention relates to an improvement in the operation efficiency of an apparatus that circulates refrigerant by a vapor compression refrigeration cycle and performs efficient air conditioning, for example, an air conditioner that can adjust indoor temperature and humidity.

従来のこの種の空気調和機、すなわち温度と湿度双方を制御する空気調和機においては、容量の異なる独立した２つの冷凍サイクルで異なる蒸発温度を発生させ、温湿度を制御するものがある（例えば、特許文献１参照）。また、冷凍サイクルによる冷房と、その排熱を利用した除湿機により温湿度制御を行い、それらを連係制御することで空調システムの運転効率を向上する技術が知られている（例えば、特許文献２参照）。更には外気導入を行うことにより室内に設けた顕熱冷房ユニットの配置が自由に出来る技術が知られている。（例えば、特許文献３参照）   In this type of conventional air conditioner, that is, an air conditioner that controls both temperature and humidity, there are those that generate different evaporating temperatures in two independent refrigeration cycles having different capacities to control temperature and humidity (for example, , See Patent Document 1). Further, a technique is known in which temperature and humidity control is performed by cooling using a refrigeration cycle and a dehumidifier using the exhaust heat, and the operation efficiency of the air conditioning system is improved by linking and controlling them (for example, Patent Document 2). reference). Further, a technique is known in which the arrangement of a sensible heat cooling unit provided indoors can be freely performed by introducing outside air. (For example, see Patent Document 3)

特開平１０−２５９９４４号公報（請求項１、０１００欄、図１、図２等）JP-A-10-259944 (Claim 1, 0100 column, FIG. 1, FIG. 2, etc.) 特開２００２−２２２４５号公報（００３３欄、図１等）Japanese Patent Laid-Open No. 2002-22245 (column 0033, FIG. 1, etc.) 特開２００５−４９０５９号公報（請求項１、図２等）Japanese Patent Laying-Open No. 2005-49059 (Claim 1, FIG. 2, etc.)

しかしながら、同一室内に複数系統の冷凍サイクルを配備する場合、冷媒配管工事コストが増大するという問題がある。また、顕熱比の大きい運転、すなわち高い蒸発温度で運転する側の冷凍サイクルでは圧縮比が小さくなり効率のよい運転が可能となるが、顕熱比の小さい運転、すなわち低い蒸発温度で運転する側の冷凍サイクルでは圧縮比が大きくなり、消費電力の大きい運転となってしまう。 However, when a plurality of refrigeration cycles are installed in the same room, there is a problem that the cost of refrigerant piping construction increases. In addition, in a refrigeration cycle that operates at a high sensible heat ratio, that is, a refrigeration cycle that operates at a high evaporation temperature, an efficient operation is possible, but an operation at a low sensible heat ratio, that is, an operation at a low evaporation temperature. In the refrigeration cycle on the side, the compression ratio becomes large and the operation consumes a large amount of power.

また、冷凍サイクルの排熱を利用した除湿機を備えた空調システムで成績係数（ＣＯＰ）を最大化するように冷凍サイクル運転制御を行う場合においても、室内ユニットへの冷媒配管以外に除湿ユニットへの冷媒配管が追加で必要となるため、工事コストが増大するという問題がある。またおのおののユニットで最適な運転を行っても消費電力の低減が行えないという問題がある。 In addition, when the refrigeration cycle operation control is performed so as to maximize the coefficient of performance (COP) in an air conditioning system equipped with a dehumidifier utilizing exhaust heat of the refrigeration cycle, the dehumidification unit is not limited to the refrigerant pipe to the indoor unit. However, there is a problem that the construction cost increases. There is also a problem that power consumption cannot be reduced even if each unit is optimally operated.

この発明の目的は、上記のような課題を解決するためになされたもので、本発明は各々のユニットを接続する冷媒配管を１系統として工事負荷を軽減することが出来、所定の冷房負荷に対して最大効率で運転することが可能な空気調和機を得ることを目的とする。また本発明は環境問題を改善できる簡単な構造の効率の良い空気調和機を得ることを目的とする。   The object of the present invention is to solve the above-mentioned problems, and the present invention can reduce the work load by using one refrigerant pipe connecting each unit as a system, so that a predetermined cooling load can be achieved. On the other hand, an object is to obtain an air conditioner capable of operating at maximum efficiency. Another object of the present invention is to obtain an efficient air conditioner having a simple structure that can improve environmental problems.

この発明に係る空気調和機は、冷媒を吐出する第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し冷媒を循環させる第１の冷媒サイクルと、この第１の冷媒サイクルに第１の減圧手段および第１の負荷側熱交換器と並列に接続される第２の減圧手段、第２の負荷側熱交換器および第１の圧縮機とは独立に運転される第２の圧縮機を有する第２の冷媒サイクルと、第１の負荷側熱交換器及び第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、第１の負荷側熱交換器の表面温度が空調領域の空気の露点温度より高くなり、第２の負荷側熱交換器の表面温度が空調領域の露点温度より低くなる様に、第１の負荷側熱交換器の熱処理能力を第２の負荷側熱交換器の熱処理能力より大きくしたものである。   In the air conditioner according to the present invention, a first compressor that discharges refrigerant, a heat source side heat exchanger, a first pressure reducing unit, and a first load side heat exchanger are sequentially connected to circulate the refrigerant. A refrigerant cycle, a second decompression means connected in parallel with the first decompression means and the first load-side heat exchanger, a second load-side heat exchanger and a first compression in the first refrigerant cycle The air conditioned by the second refrigerant cycle having the second compressor that is operated independently of the compressor, the first load side heat exchanger, and the second load side heat exchange in the same air conditioning region A load-side air blowing means for blowing out, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is dew point of the air-conditioning region The heat treatment capacity of the first load-side heat exchanger is adjusted to the heat treatment of the second load-side heat exchanger so as to be lower than the temperature. It is made larger than the force.

この発明に係る空気調和機は、容量調節可能な第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し冷媒を循環させる第１の冷媒サイクルと、第１の減圧手段を形成すると共にこの第１の減圧手段と並列に第２の減圧手段および第２の負荷側熱交換器を接続し、第１の減圧手段の減圧時の駆動流による吸引を利用して第２の負荷側熱交換器からの冷媒圧力を昇圧する昇圧手段と、第１の負荷側熱交換器及び第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、第１の負荷側熱交換器の表面温度が空調領域の空気の露点温度より高くなり、第２の負荷側熱交換器の表面温度が空調領域の露点温度より低くなる様に、第１の負荷側熱交換器の熱処理能力を第２の負荷側熱交換器の熱処理能力より大きくしたものである。   The air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, a first pressure reducing unit, and a first load side heat exchanger, which are sequentially connected to circulate the refrigerant. A refrigerant cycle and a first decompression unit are formed, and a second decompression unit and a second load-side heat exchanger are connected in parallel with the first decompression unit, and the first decompression unit is driven during decompression. The pressure-increasing means for increasing the refrigerant pressure from the second load-side heat exchanger using suction by flow is the same as the air conditioned by the first load-side heat exchanger and the second load-side heat exchange Load-side air blowing means that blows out to the air-conditioning region, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is The heat treatment capacity of the first load-side heat exchanger is set so that it is lower than the dew point temperature of the air-conditioning area. It is made larger than the heat treatment capacity of the exchanger.

この発明に係る空気調和機は、容量調節可能な第１の圧縮機、熱源側熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、第２の減圧手段、第２の室内熱交換器、風量調節可能な第２の室内送風機、および第１の圧縮機とは独立に容量調節可能な第２の圧縮機、を具備する第２の室内ユニットと、室外ユニット、第１の室内ユニット、第２の室内ユニットを接続する冷媒配管と、を備え、第１の室内熱交換器表面温度が室内空気の露点温度より高くなるように運転されるものである。   An air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, a first The capacity can be adjusted independently of the first indoor unit including the indoor blower, the second decompression means, the second indoor heat exchanger, the second indoor blower with adjustable air volume, and the first compressor. A second indoor unit comprising a second compressor, an outdoor unit, a first indoor unit, and a refrigerant pipe connecting the second indoor unit, and a first indoor heat exchanger surface temperature Is operated so as to be higher than the dew point temperature of room air.

この発明に係る空気調和機は、容量調節可能な第１の圧縮機、熱源側熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、膨張動力回収手段、第２の減圧手段、第２の室内熱交換器、第２の室内送風機、第２の圧縮機を具備する第２の室内ユニットと、室外ユニット、第１の室内ユニット、第２の室内ユニットを高圧側液配管と低圧側ガス配管で接続し、高圧側の冷媒により膨張動力回収手段で回収された動力で第２の圧縮機を駆動するものである。   An air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, a first A second indoor unit including a first indoor unit including an indoor fan, an expansion power recovery unit, a second decompression unit, a second indoor heat exchanger, a second indoor fan, and a second compressor. The unit, the outdoor unit, the first indoor unit, and the second indoor unit are connected by the high pressure side liquid pipe and the low pressure side gas pipe, and the second compression is performed by the power recovered by the expansion power recovery means by the high pressure side refrigerant. It drives the machine.

この発明の空気調和機は、顕熱と潜熱を処理して空気調和を行う熱交換器を別々に設けたので簡単な構造で効率の良い運転が可能になる。又本発明は低段と高段の異なる２つの蒸発圧力で同一の冷房負荷を処理する際に、低段側蒸発温度を消費電力が最小となるようにしたので、高効率な運転を行うことができる。   The air conditioner according to the present invention is provided with a separate heat exchanger for processing sensible heat and latent heat to perform air conditioning, so that it is possible to operate efficiently with a simple structure. In the present invention, when the same cooling load is processed with two different evaporating pressures of the low stage and the high stage, the low stage side evaporating temperature is minimized so that the power consumption is minimized. Can do.

また、この発明は、高段側熱交換器で空気を予冷した後に低段側熱交換器で冷却除湿するので、低段側の冷媒流量比が高段側に対して小さくなり、トータルの消費電力を小さくすることができる。また、この発明は目的や特性に応じた効率の良い熱交換器の構造や配置が可能である。   Further, according to the present invention, since the air is pre-cooled by the high-stage heat exchanger and then cooled and dehumidified by the low-stage heat exchanger, the refrigerant flow ratio on the low-stage side is smaller than that on the high-stage side, and the total consumption is reduced. Electric power can be reduced. In addition, the present invention enables an efficient heat exchanger structure and arrangement according to the purpose and characteristics.

実施の形態１．
図１はこの発明の実施の形態における空気調和機の冷媒回路の一例を示すものである。図１において、室外などに設けられ熱源側である室外ユニット１、室内壁面などに配置され負荷側である第１の室内ユニット２、同様に負荷側であるが少なくとも吹出口は室内側に設けられた第２の室内ユニット３である。第１および第２の室内ユニット２、３は、空調領域である同一空間の空調負荷を処理するように配置されており、また、図においては各１台であるがそれぞれ複数台であってもよい。封入されている冷媒はフロン系冷媒であるＲ４１０Ａ、もしくは自然冷媒、例えば炭酸ガス、イソブタン、水などである。なお図１は室内ユニットにて冷房を行う方向で冷媒を循環させる構成を説明しているが、室外ユニット１、第２の室内ユニット３に設けたそれぞれの圧縮機４、１４の吐出口に四方弁を設け吐出方向を切換えて循環する冷媒回路を逆にすることにより暖房も可能である冷凍サイクルに変更できることは言うまでもない。又第１の圧縮機はインバータ駆動で容量を調整できるものとして説明するが、常に最大容量で運転を継続させる回転速度を調節しない誘導電動機等と制御を使用し、圧縮機をオンとオフにより調節しても良いし、これにより効率の良い運転と、簡単な構造の装置が得られる。 Embodiment 1 FIG.
FIG. 1 shows an example of a refrigerant circuit of an air conditioner according to an embodiment of the present invention. In FIG. 1, an outdoor unit 1 provided on the outdoor side or the like, which is on the heat source side, a first indoor unit 2 which is disposed on the indoor wall surface or the like and is on the load side, and similarly on the load side, but at least the outlet is provided on the indoor side. This is the second indoor unit 3. The first and second indoor units 2 and 3 are arranged so as to process the air conditioning load in the same space as the air conditioning area. Good. The encapsulated refrigerant is R410A, which is a fluorocarbon refrigerant, or a natural refrigerant such as carbon dioxide, isobutane, or water. FIG. 1 illustrates the configuration in which the refrigerant is circulated in the direction of cooling in the indoor unit, but there are four directions at the discharge ports of the compressors 4 and 14 provided in the outdoor unit 1 and the second indoor unit 3, respectively. It goes without saying that the refrigerant circuit can be changed to a refrigeration cycle that can be heated by providing a valve and switching the discharge direction to reverse the circulating refrigerant circuit. The first compressor is explained as being capable of adjusting the capacity by driving the inverter. However, the compressor is controlled by turning the compressor on and off using an induction motor and the like that do not adjust the rotational speed that always keeps operating at the maximum capacity. Alternatively, an efficient operation and a device having a simple structure can be obtained.

室外ユニット１は、運転容量調節可能なスクロールやロータリーなどの冷媒を吸込み吐出する高段圧縮機４、室外空気と熱交換する室外熱交換器５、その熱交換量を調節する室外送風機６で構成される。７、８は冷媒配管であり、室外ユニット１と室内ユニット２、３を接続している。高段圧縮機４から吐出された高温、高圧ガス冷媒は室外熱交換器５にて凝縮し室外送風機６が送風する室外空気と熱交換し高圧液冷媒となり液配管７にて室内へ運ばれる。   The outdoor unit 1 includes a high-stage compressor 4 that sucks and discharges a refrigerant such as a scroll or a rotary that can adjust its operating capacity, an outdoor heat exchanger 5 that exchanges heat with outdoor air, and an outdoor fan 6 that adjusts the amount of heat exchange. Is done. Reference numerals 7 and 8 denote refrigerant pipes that connect the outdoor unit 1 and the indoor units 2 and 3. The high-temperature and high-pressure gas refrigerant discharged from the high-stage compressor 4 condenses in the outdoor heat exchanger 5, exchanges heat with the outdoor air blown by the outdoor blower 6, becomes high-pressure liquid refrigerant, and is carried indoors through the liquid pipe 7.

第１の室内ユニット２は、第１減圧装置９、第１室内熱交換器１０、第１室内送風機１１により構成されている。この第１の室内ユニット２はその熱交換量に対して十分大きい熱交換能力である伝熱面積と送風量を有しており、第１室内熱交換器１０を流通する冷媒温度は室内空気の露点温度以上となる。よって、この第１室内ユニット２において室内空気は除湿されず、顕熱負荷のみが処理される。ここで第１の室内ユニット２の第１室内熱交換器と第１室内送風機が、その熱交換量に対して十分大きい伝熱性能（伝熱面積と送風量）を有しているとは、例えばその内部を通過する冷媒と第１室内送風機が送風する室内空気と熱交換を行うが、冷媒と室内空気の温度差が１０度以内となる程度の伝熱面積と送風量に設定していることである。例えば室内空気の吸込み温度を２７℃とし湿度５０％、露点温度１５℃としたときに、吸込み空気温度−蒸発温度が１０度以下程度とすることにより、冷房時に除湿がほとんど行われず高い蒸発温度となり圧縮比が小さく効率の高い運転が可能になる。上記のように熱交換器の伝熱性能は熱交換する伝熱面積と風量で決まり、空気温度と蒸発温度差が１０度程度以内となる伝熱性能を有していれば、ほぼ蒸発温度は露点温度以上となる。なお通常冷媒温度は熱交換器の伝熱管表面に設けた管温温度計で計測している。   The first indoor unit 2 includes a first pressure reducing device 9, a first indoor heat exchanger 10, and a first indoor blower 11. The first indoor unit 2 has a heat transfer area and an air blowing amount that are sufficiently large in heat exchange capacity with respect to the heat exchange amount, and the temperature of the refrigerant flowing through the first indoor heat exchanger 10 is the indoor air. Above dew point temperature. Therefore, in this 1st indoor unit 2, indoor air is not dehumidified but only a sensible heat load is processed. Here, the first indoor heat exchanger and the first indoor fan of the first indoor unit 2 have sufficiently large heat transfer performance (heat transfer area and air flow rate) with respect to the heat exchange amount. For example, heat exchange is performed between the refrigerant passing through the interior and the indoor air blown by the first indoor blower, but the heat transfer area and the air flow rate are set so that the temperature difference between the refrigerant and the indoor air is within 10 degrees. That is. For example, when the intake temperature of indoor air is 27 ° C., the humidity is 50%, and the dew point temperature is 15 ° C., the intake air temperature-evaporation temperature is about 10 ° C. or less, so that dehumidification is hardly performed during cooling and the evaporation temperature is high. Operation with a low compression ratio and high efficiency is possible. As described above, the heat transfer performance of the heat exchanger is determined by the heat transfer area and the air volume for heat exchange, and if the heat transfer performance is such that the difference between the air temperature and the evaporation temperature is within about 10 degrees, the evaporation temperature is almost Above dew point temperature. The refrigerant temperature is usually measured by a tube temperature thermometer provided on the heat transfer tube surface of the heat exchanger.

第１室内ユニットで使用される顕熱熱交換器は露点温度以上となるため熱交換器の表面にほとんど結露することが無く、ドレンパン、ドレンポンプ、ドレン配管などの結露水処理機構が不要であるか、或いは底の浅いドレンパンのみを設け若干溜まった結露水を自然蒸発させる程度の簡単な構造とすることが出来る。したがって第１室内ユニットの構成や第１の負荷側熱交換器周りが簡素化され機器コストや工事コストを低減することが出来る。この様に高段側では室内空気の目標とする設定温度と計測する実際の室内空気温度の差で第１の圧縮機を制御することになる。このため室内が異常に高湿度である場合は第１の負荷側熱交換器である顕熱熱交換器にも結露を発生する可能性が残るので熱交換器下部には貯水量の少ない構造のドレンパンを設けたり、熱交換器構造でも冷却フィンを通して若干の結露水の流下を可能にする簡単な流路を設けたりする。   Since the sensible heat exchanger used in the first indoor unit has a dew point temperature or higher, there is almost no condensation on the surface of the heat exchanger, and there is no need for a dew condensation water treatment mechanism such as a drain pan, drain pump, drain pipe, etc. Alternatively, it is possible to provide a simple structure in which only a drain pan having a shallow bottom is provided and the condensed water slightly accumulated is naturally evaporated. Therefore, the configuration of the first indoor unit and the surroundings of the first load side heat exchanger are simplified, and the equipment cost and the construction cost can be reduced. In this way, on the high stage side, the first compressor is controlled by the difference between the target set temperature of the room air and the actual room air temperature to be measured. For this reason, when the room is abnormally high in humidity, the sensible heat exchanger, which is the first load-side heat exchanger, has the possibility of causing condensation. A drain pan is provided, or a simple flow path that allows some condensed water to flow through the cooling fins even in the heat exchanger structure.

次に第２の室内ユニット３は、第２減圧装置１２、第２室内熱交換器１３、低段圧縮機１４、第２室内送風機１５によって構成され液配管７とガス配管８に第１の室内ユニットと並列に接続されており、ここでの冷媒温度は室内空気の露点温度以下となるように運転される。よって、この第２の室内ユニット３では顕熱負荷と潜熱負荷双方が処理され、室外機を凝縮運転する冷房運転時には室内空気の冷却とともに除湿が行われる。このため第２の負荷側熱交換器に対しては結露水対策が必要である。以上のように本発明の冷凍サイクルでは室外ユニット１で凝縮運転したときに２つの室内ユニットでは膨張機９、１２の開度設定を変えて異なる蒸発温度の蒸発運転を行うが、この冷凍サイクルを空気調和機に使用する場合、高い蒸発温度のほうでは面積の大きな熱交換器を使用して顕熱主体の運転とし、低い蒸発温度のほうでは潜熱主体運転のように送風手段の風量、第２の減圧手段の開度、などにより処理熱量のコントロールを行えるようにして、高効率での冷房運転や除湿運転を行う。また、各圧縮機と組み合わせる四方弁を設け暖房運転、すなわち各室内ユニットではそれぞれ凝縮運転を行うことも出来る。   Next, the second indoor unit 3 includes a second decompression device 12, a second indoor heat exchanger 13, a low-stage compressor 14, and a second indoor blower 15, and is connected to the liquid pipe 7 and the gas pipe 8 in the first room. It is connected in parallel with the unit, and the refrigerant temperature here is operated so as to be equal to or lower than the dew point temperature of the room air. Therefore, both the sensible heat load and the latent heat load are processed in the second indoor unit 3, and dehumidification is performed along with cooling of the indoor air during the cooling operation in which the outdoor unit is condensed. For this reason, a countermeasure against condensed water is necessary for the second load-side heat exchanger. As described above, in the refrigeration cycle of the present invention, when the outdoor unit 1 performs the condensation operation, the two indoor units perform the evaporation operation at different evaporation temperatures by changing the opening settings of the expanders 9 and 12. When used for an air conditioner, the higher evaporation temperature uses a heat exchanger with a larger area to operate mainly by sensible heat, and the lower evaporation temperature uses the air volume of the blowing means as in the latent heat main operation. The amount of heat treated can be controlled by the opening degree of the decompression means, etc., so that the cooling operation and the dehumidification operation are performed with high efficiency. Further, a four-way valve combined with each compressor can be provided to perform heating operation, that is, each indoor unit can perform condensation operation.

次に、このように構成された本実施の形態１の空気調和機の冷凍サイクル動作について、図１および図２を参照して説明する。図２は本実施の形態１の空気調和機による冷凍サイクル動作を示すＰ−ｈ線図で、横軸は比エンタルピー［ｋＪ／ｋｇ］、縦軸は冷媒圧力［ＭＰａ］である。また、図２中に矢印で示したＧ１、Ｇ２はそれぞれ冷媒流量［ｋｇ／ｈ］を示している。   Next, the refrigeration cycle operation of the air conditioner of the first embodiment configured as described above will be described with reference to FIG. 1 and FIG. FIG. 2 is a Ph diagram illustrating the refrigeration cycle operation by the air conditioner of the first embodiment. The horizontal axis represents specific enthalpy [kJ / kg], and the vertical axis represents refrigerant pressure [MPa]. Further, G1 and G2 indicated by arrows in FIG. 2 respectively indicate the refrigerant flow rate [kg / h].

高段圧縮機４から吐出された高温高圧のガス冷媒（状態Ａ）は、室外熱交換器５において室外送風機６により外気に放熱して凝縮し、高圧液冷媒（状態Ｂ）となる。この高圧液冷媒は冷媒配管７を流通、第１減圧装置９により減圧されて飽和温度で約２０℃の中間圧力Ｐｅ１（状態Ｃ）まで減圧され、第１室内熱交換器１０により室内空気と熱交換を行い蒸発する。一方、第２減圧装置１２では飽和温度で約５℃の低圧Ｐｅ２（状態Ｅ）まで減圧され、第２室内熱交換器１３により室内空気と熱交換を行い蒸発する。この低圧ガス冷媒（状態Ｆ）は低段圧縮機１４により中圧Ｐｅ１まで昇圧され（状態Ｇ）、第１室内熱交換器１０を流出した中圧ガス冷媒（状態Ｄ）と合流して（状態Ｈ）、再び高段圧縮機４に吸入される。   The high-temperature and high-pressure gas refrigerant (state A) discharged from the high-stage compressor 4 dissipates heat to the outside air by the outdoor fan 6 in the outdoor heat exchanger 5 and condenses to become high-pressure liquid refrigerant (state B). This high-pressure liquid refrigerant flows through the refrigerant pipe 7 and is decompressed by the first decompression device 9 and decompressed to an intermediate pressure Pe1 (state C) of about 20 ° C. at the saturation temperature. Exchange and evaporate. On the other hand, in the second pressure reducing device 12, the pressure is reduced to a low pressure Pe2 (state E) of about 5 ° C. at the saturation temperature, and the second indoor heat exchanger 13 exchanges heat with room air to evaporate. This low-pressure gas refrigerant (state F) is pressurized to the intermediate pressure Pe1 by the low-stage compressor 14 (state G), and merges with the medium-pressure gas refrigerant (state D) that has flowed out of the first indoor heat exchanger 10 (state D). H), and again sucked into the high stage compressor 4.

続いて、この図２に示した本実施の形態１における冷凍サイクルの成績係数（ＣＯＰ）について説明する。ＣＯＰとはこの冷凍サイクルの冷却能力を圧縮機入力で除したもので、その値が大きいほど高効率な冷凍サイクルであるといえる。   Next, the coefficient of performance (COP) of the refrigeration cycle in the first embodiment shown in FIG. 2 will be described. The COP is obtained by dividing the cooling capacity of the refrigeration cycle by the compressor input, and the higher the value, the higher the efficiency of the refrigeration cycle.

図２中に矢印で示したように、第１室内熱交換器を流通する冷媒流量をＧ１、第２室内熱交換器を流通する冷媒流量をＧ２とすると、低段圧縮機１４では冷媒流量Ｇ２が昇圧され、高段圧縮機４では合流後の冷媒流量（Ｇ１＋Ｇ２）が昇圧される。Δｈ１、Δｈ２を、高段側、低段側それぞれの単位冷媒流量あたりの圧縮動力とすると、低段圧縮機１４の入力はＧ２×Δｈ２、高段圧縮機４の入力は（Ｇ１＋Ｇ２）×Δｈ１で与えられる。 As indicated by an arrow in FIG. 2, when the refrigerant flow rate flowing through the first indoor heat exchanger is G1, and the refrigerant flow rate flowing through the second indoor heat exchanger is G2, the low-stage compressor 14 uses the refrigerant flow rate G2. And the refrigerant flow rate (G1 + G2) after merging is increased in the high stage compressor 4. If Δh1 and Δh2 are compression power per unit refrigerant flow rate on the high stage side and the low stage side, the input of the low stage compressor 14 is G2 × Δh2, and the input of the high stage compressor 4 is (G1 + G2) × Δh1. Given.

この冷凍サイクルの冷却能力は蒸発器エンタルピ差Δｈｅと冷媒流量との積で求められる。ここで、冷媒をＨＦＣとすると第１室内熱交換器と第２室内熱交換器それぞれの出口エンタルピ（状態Ｄ、状態Ｆ）はほとんど等しいため、それぞれの熱交換器での蒸発エンタルピ差Δｈｅが等しいとすると、冷却能力は（Ｇ１＋Ｇ２）×Δｈｅとなる。よってＣＯＰは式（１）により与えられる。
ＣＯＰ＝（Ｇ１＋Ｇ２）Δｈｅ／（（Ｇ１＋Ｇ２）Δｈ１＋Ｇ２Δｈ２）…式（１） The cooling capacity of this refrigeration cycle is obtained by the product of the evaporator enthalpy difference Δhe and the refrigerant flow rate. Here, if the refrigerant is HFC, the outlet enthalpies (state D and state F) of the first indoor heat exchanger and the second indoor heat exchanger are almost equal, so the evaporation enthalpy difference Δhe in each heat exchanger is equal. Then, the cooling capacity is (G1 + G2) × Δhe. Thus, COP is given by equation (1).
COP = (G1 + G2) Δhe / ((G1 + G2) Δh1 + G2Δh2) (1)

この式（１）は、この空気調和機の冷却能力（Ｇ１＋Ｇ２）Δｈｅが一定のとき、高段圧縮機４の入力はＧ１とＧ２の比によらずΔｈ１のみにより決定され、また、低段圧縮機入力は、Ｇ２が小さいほど、そしてΔｈ２が小さいほど小さくなることを示している。よって、この冷凍サイクルのＣＯＰを大きくするには、Δｈ１を小さくする、すなわち、Ｐｅ１を上昇させること、そして、Δｈ２を小さくすること、そして（Ｇ１＋Ｇ２）一定のもとでＧ２を小さくすることといえる。ただし、高段蒸発圧力Ｐｅ１は第１室内熱交換器１０の伝熱性能で決まるため、ある値以上に上昇させることは不可能である。   This equation (1) indicates that when the cooling capacity (G1 + G2) Δhe of this air conditioner is constant, the input of the high-stage compressor 4 is determined only by Δh1 regardless of the ratio of G1 and G2, and the low-stage compression The machine input shows that the smaller G2 and the smaller Δh2, the smaller. Therefore, to increase the COP of this refrigeration cycle, it can be said that Δh1 is decreased, that is, Pe1 is increased, Δh2 is decreased, and G2 is decreased under a constant (G1 + G2). . However, since the high stage evaporation pressure Pe1 is determined by the heat transfer performance of the first indoor heat exchanger 10, it cannot be increased to a certain value or more.

一方で、低段側の第２室内熱交換器での除湿特性は、冷媒の蒸発圧力である蒸発温度（Ｐｅ２）が上昇するのに伴い、顕熱比（全冷却能力に対する顕熱能力の割合）が上昇するため、同じ除湿量を得るための冷媒流量Ｇ２は大きくなる。図３は低段熱交換器の蒸発温度変化に対する顕熱比特性を示す図で、横軸に蒸発温度、縦軸に顕熱比をとったグラフで、蒸発温度が変化した場合の顕熱比特性を吸込温度２７℃、吸込相対湿度５０％、６０％、７０％の３通りについて示しているが、この図に示すように、蒸発温度が吸込空気の露点温度に漸近するに従い、顕熱比は急激に上昇する。よって、あるところまでは蒸発圧力Ｐｅ２を高くすることでΔｈ２減少によるＣＯＰ向上効果が得られるが、所定値以上に蒸発温度が上昇すると、同じ除湿量を得るための低段側冷媒流量Ｇ２の急増によってＣＯＰが悪化する。図４は低段蒸発温度に対する低段圧縮機入力変化を示す図で、除湿量一定条件での冷媒流量、単位流量当りの圧縮動力、圧縮機入力の変化を、横軸を蒸発温度として示す。   On the other hand, the dehumidification characteristic in the second indoor heat exchanger on the lower stage side is the sensible heat ratio (the ratio of the sensible heat capacity to the total cooling capacity) as the evaporation temperature (Pe2) that is the evaporation pressure of the refrigerant increases. ) Increases, the refrigerant flow rate G2 for obtaining the same dehumidification amount increases. FIG. 3 is a graph showing the sensible heat ratio characteristics with respect to the evaporation temperature change of the low stage heat exchanger. The horizontal axis represents the evaporation temperature and the vertical axis represents the sensible heat ratio. The characteristics are shown for three types of suction temperature 27 ° C., suction relative humidity 50%, 60%, and 70%. As shown in this figure, as the evaporation temperature gradually approaches the dew point temperature of the suction air, the sensible heat ratio Rises rapidly. Therefore, by increasing the evaporation pressure Pe2 up to a certain point, a COP improvement effect can be obtained by reducing Δh2, but when the evaporation temperature rises above a predetermined value, the low-stage refrigerant flow rate G2 increases rapidly to obtain the same dehumidification amount. COP deteriorates. FIG. 4 is a diagram showing a change in the low-stage compressor input with respect to the low-stage evaporation temperature, and shows the change in refrigerant flow rate, compression power per unit flow rate, and compressor input under the constant dehumidification amount as the evaporation temperature on the horizontal axis.

図４に示したように、蒸発圧力Ｐｅ２の上昇はΔｈ２を減少させるというＣＯＰを向上する方向に働く作用と、冷媒流量Ｇ２を増大させるというＣＯＰを悪化させる作用双方をもち、低段圧縮機入力を最小とする、言い換えるとＣＯＰを最大とする最適蒸発温度が存在することがわかる。ただし、ここでは第１の室内熱交換器１０は十分大きな伝熱性能を有しており、かつ、高段側冷媒流量Ｇ１は低段側冷媒流量Ｇ２に対して十分大きく、低段側冷媒流量Ｇ２の変化に対して高段側蒸発圧力Ｐｅ１はほとんど影響を受けないことを前提としている。   As shown in FIG. 4, the increase in the evaporation pressure Pe2 has both the effect of improving the COP in which Δh2 is reduced and the effect of deteriorating the COP in which the refrigerant flow rate G2 is increased. It can be seen that there exists an optimum evaporating temperature that minimizes COP, in other words, that maximizes COP. However, here, the first indoor heat exchanger 10 has sufficiently large heat transfer performance, and the high-stage refrigerant flow rate G1 is sufficiently larger than the low-stage refrigerant flow rate G2, and the low-stage refrigerant flow rate is low. It is assumed that the high-stage evaporation pressure Pe1 is hardly affected by changes in G2.

上述のように、本発明の実施の形態における空気調和機では、低段圧縮機入力が最も小さくなる最適蒸発温度を推定し、実際の蒸発温度がその値に近づくように運転制御が行われる。続いて、この発明の実施の形態における空気調和機の制御の一例について図５の制御フローチャートを参照して説明する。なお実際の蒸発温度とは蒸発器である各室内熱交換器の冷媒が流れる管の温度を直接計測し、この管の外側で計測した温度を蒸発温度とする。   As described above, in the air conditioner according to the embodiment of the present invention, the optimum evaporation temperature at which the low-stage compressor input is minimized is estimated, and operation control is performed so that the actual evaporation temperature approaches that value. Next, an example of control of the air conditioner in the embodiment of the present invention will be described with reference to the control flowchart of FIG. The actual evaporating temperature directly measures the temperature of the pipe through which the refrigerant of each indoor heat exchanger that is an evaporator flows, and the temperature measured outside this pipe is taken as the evaporating temperature.

図示は省略するが、室内には温度センサおよび湿度センサが設置され、ステップＳ１で随時室内空気状態（温度Ｔと湿度ＲＨ）が検知される。また、リモコン等のユーザーインターフェースにより目標となる設定温度Ｔ*、設定湿度ＲＨ*が設定されている。さらに、現在の冷凍サイクル動作状態として高段、低段の蒸発温度が検知される。（Ｓ1ａ、Ｓ1ｂ、Ｓ1ｃ）   Although illustration is omitted, a temperature sensor and a humidity sensor are installed in the room, and an indoor air condition (temperature T and humidity RH) is detected at any time in step S1. Also, a target set temperature T * and set humidity RH * are set by a user interface such as a remote controller. Furthermore, high and low evaporating temperatures are detected as the current refrigeration cycle operating state. (S1a, S1b, S1c)

ステップＳ２では、設定温湿度と現在の温湿度から顕熱負荷、潜熱負荷がどの程度かを推定する。湿度偏差ΔＲＨは潜熱負荷の大きさを示し、温度偏差ΔＴは顕熱負荷の大きさを示すと考える。前述したように、第１室内熱交換器１０は顕熱負荷のみを処理し、第２室内熱交換器１３は顕熱負荷、潜熱負荷双方を処理するので、潜熱負荷の大小にしたがって低段圧縮機１４の容量制御、顕熱負荷の大小にしたがって高段圧縮機４の容量を制御するのが妥当である。   In step S2, the sensible heat load and the latent heat load are estimated from the set temperature and humidity and the current temperature and humidity. It is assumed that the humidity deviation ΔRH indicates the magnitude of the latent heat load, and the temperature deviation ΔT indicates the magnitude of the sensible heat load. As described above, the first indoor heat exchanger 10 processes only the sensible heat load, and the second indoor heat exchanger 13 processes both the sensible heat load and the latent heat load. It is appropriate to control the capacity of the high stage compressor 4 according to the capacity control of the machine 14 and the magnitude of the sensible heat load.

ステップＳ３では、第２室内熱交換器１３の最適蒸発温度Ｔｅ２*を計算する。このＴｅ２*は、現在の室内空気状態および潜熱負荷の大きさにより求められる値であり、図６に示すような関係から求める。図６は低段蒸発温度とＣＯＰとの関係を示す図であって、これには室内空気が２７℃５０％のときの蒸発温度とＣＯＰとの関係を示しており、図１に示す冷凍サイクル全体のＣＯＰと低段蒸発温度との関係を示す。このような特性を発生し得る空気条件、すなわち熱負荷、温度、湿度の関係を空気線図で把握しておくことによってＣＯＰ最大となる蒸発温度Ｔｅ２*を演算することができる。なおＳＨＦは全体の処理すべき冷房負荷量に対する顕熱負荷量との比である顕熱比を示し、このＳＨＦ毎の蒸発温度が変化するときのＣＯＰ最大となる蒸発温度Ｔｅ２*を演算することにより最も効率の良い目標蒸発温度を探すものである。ＳＨＦの代わりに潜熱負荷量をパラメータにしても良い。例えばＳＨＦ０．８５とは１５％の潜熱負荷が発生していることを示し、ＳＨＦが大きくなるにつれてＣＯＰの絶対値が上昇しているのは潜熱負荷が小さくなることにより低段冷媒流量Ｇ２が減少しているためである。   In step S3, the optimal evaporation temperature Te2 * of the second indoor heat exchanger 13 is calculated. This Te2 * is a value obtained from the current indoor air condition and the magnitude of the latent heat load, and is obtained from the relationship shown in FIG. FIG. 6 is a diagram showing the relationship between the low-stage evaporation temperature and the COP. This shows the relationship between the evaporation temperature and the COP when the room air is 27 ° C. and 50%. The refrigeration cycle shown in FIG. The relationship between the whole COP and the low stage evaporation temperature is shown. The air temperature that can generate such characteristics, that is, the relationship between the heat load, temperature, and humidity is grasped by an air diagram, and the evaporation temperature Te2 * at which the COP becomes maximum can be calculated. SHF indicates a sensible heat ratio, which is a ratio of the sensible heat load to the cooling load to be processed as a whole, and calculates the evaporation temperature Te2 * which is the maximum COP when the evaporation temperature for each SHF changes. To find the most efficient target evaporation temperature. The latent heat load may be used as a parameter instead of SHF. For example, SHF 0.85 indicates that a latent heat load of 15% is generated, and the absolute value of COP increases as SHF increases because the latent heat load decreases and the low stage refrigerant flow rate G2 decreases. It is because it is doing.

ステップＳ４では、前述のように、高段圧縮機、低段圧縮機を負荷に応じて回転数制御するとともに、第２の室内熱交換器１３での蒸発温度Ｔｅ２を最適値Ｔｅ２*となるように第２室内送風機１５の送風量を調節する。また、図示はしていないが、第１の室内熱交換器１０ではその蒸発圧力Ｐｅ１が高いほど効率のよい運転をするので、第１室内送風機はおおよそ最大風量で運転するよう制御する。又ステップ４で目標に到達すれば、例えば除湿負荷がなくなれば第２圧縮機の役割は終了し、第２冷媒サイクルの目標蒸発温度も露点温度以上になる。即ちステップ４に到達した後、熱処理はそのまま維持される。ΔＲＨやΔＴなどがゼロになると圧縮機回転数はその時の値を維持し、差分がマイナスになるとはじめて回転数が低下したり停止することになる。   In step S4, as described above, the rotational speed of the high-stage compressor and the low-stage compressor is controlled according to the load, and the evaporation temperature Te2 in the second indoor heat exchanger 13 is set to the optimum value Te2 *. The air flow rate of the second indoor blower 15 is adjusted. Although not shown, the first indoor heat exchanger 10 operates more efficiently as its evaporation pressure Pe1 is higher. Therefore, the first indoor fan is controlled to operate at approximately the maximum air flow. If the target is reached in step 4, for example, if there is no dehumidification load, the role of the second compressor is terminated, and the target evaporation temperature of the second refrigerant cycle becomes equal to or higher than the dew point temperature. That is, after reaching step 4, the heat treatment is maintained as it is. When ΔRH, ΔT, and the like become zero, the compressor rotational speed maintains the value at that time, and the rotational speed decreases or stops only when the difference becomes negative.

以上の説明では主として潜熱を処理する第２の負荷側熱交換器に対しその熱処理能力を得るために現在の湿度ＲＨを湿度計で計測して目標湿度との偏差に応じて第２の圧縮機の容量を制御するものであるが、湿度を直接計測しない方法で熱処理能力を得るものでも良い。例えば冷房時の室内湿度を一定の所定値とし室内温度だけから推定した状態として空気線図から求められた露点温度より第２の負荷側熱交換器の冷媒温度がその露点温度より低くなる様にこの熱交換器用の送風手段や第２の圧縮機を運転したり、あるいは、第２の負荷側熱交換器の冷媒温度である管温と空気温度の計測値とを比較して管温が空気温度よりほぼ１０度以上低くなる様に送風機や冷媒サイクルを制御することにより湿度センサーを持たずに潜熱処理を行うことが出来る。   In the above description, in order to obtain the heat treatment capacity of the second load side heat exchanger that mainly processes the latent heat, the current humidity RH is measured with a hygrometer and the second compressor according to the deviation from the target humidity. However, the heat treatment capability may be obtained by a method in which the humidity is not directly measured. For example, the refrigerant temperature of the second load-side heat exchanger is lower than the dew point temperature from the dew point temperature obtained from the air diagram when the room humidity during cooling is a constant predetermined value and is estimated from only the room temperature. The air temperature for the heat exchanger or the second compressor is operated, or the tube temperature, which is the refrigerant temperature of the second load-side heat exchanger, is compared with the measured value of the air temperature so that the tube temperature is air. By controlling the blower and the refrigerant cycle so as to be approximately 10 degrees or more lower than the temperature, the latent heat treatment can be performed without having a humidity sensor.

第１の減圧手段である可変膨張弁９、第２の減圧手段である可変膨張弁１２は、それぞれの一端が接続された室内熱交換器１０、および１３の出口冷媒過熱度をおおよそ５ｄｅｇ程度となるように冷媒流量を調節する。なお過熱度でなく過冷却度でも良く、圧縮機以外例えば昇圧させるエジェクターのようなものでもニードル弁による開閉で制御できる。   The variable expansion valve 9 serving as the first pressure reducing means and the variable expansion valve 12 serving as the second pressure reducing means have an outlet refrigerant superheat degree of the indoor heat exchangers 10 and 13 to which the respective one ends are connected are approximately 5 deg. The refrigerant flow rate is adjusted so that Note that the degree of supercooling may be used instead of the degree of superheat, and other than the compressor, such as an ejector for increasing the pressure, can be controlled by opening and closing with a needle valve.

以上のように、この発明に関わる空気調和機は、容量調節可能な第１の圧縮機と、室外熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第２の室内送風機を具備する第１の室内ユニットと、第２の減圧手段、第２の室内熱交換器、風量調節可能な第２の室内送風機、容量調節可能な第２の圧縮機、を備えた第２の室内ユニットと、それらを接続する冷媒配管と、室内温湿度情報検出手段と、第２の室内熱交換器を流通する冷媒温度検出手段と、目標冷媒温度設定手段を備え、冷媒温度が目標冷媒温度となるように第２の室内送風機の送風量等を制御するので、所定の冷房負荷を処理するための消費電力が最小となるように運転できる。なお、圧縮機の回転数制御装置、膨張弁の開度制御装置、目標冷媒温度設定手段等のような制御装置は室外機側や室内機側など必要なところに設けおのおの個別に制御したり、１箇所にて通信手段にて纏めて制御すれば良い。   As described above, the air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, an outdoor heat exchanger, an outdoor unit including an outdoor fan, a first decompression unit, and a first indoor heat. Exchanger, first indoor unit including second indoor blower, second decompression means, second indoor heat exchanger, second indoor blower with adjustable air volume, second compression with adjustable capacity A second indoor unit provided with an air conditioner, a refrigerant pipe connecting them, an indoor temperature / humidity information detecting means, a refrigerant temperature detecting means flowing through the second indoor heat exchanger, and a target refrigerant temperature setting means. And the air flow rate of the second indoor blower is controlled so that the refrigerant temperature becomes the target refrigerant temperature, so that the power consumption for processing a predetermined cooling load can be minimized. Control devices such as a compressor rotation speed control device, an expansion valve opening control device, a target refrigerant temperature setting means, etc. are provided in necessary places such as the outdoor unit side or the indoor unit side, and are individually controlled, What is necessary is just to control collectively by a communication means in one place.

図７は冷媒回路図であって、低段冷媒流量Ｇ２を小さくする本発明の他の冷媒回路を示している。図１に示した冷媒回路と異なるところは、第２の室内ユニット３の中に、第３の減圧手段１６と第３の室内熱交換器１７を配している。図７の通り、この第３の室内熱交換器１７は、第１の室内熱交換器１０と同じ動作を行う位置に配されている。   FIG. 7 is a refrigerant circuit diagram and shows another refrigerant circuit of the present invention for reducing the low-stage refrigerant flow rate G2. A difference from the refrigerant circuit shown in FIG. 1 is that a third decompression means 16 and a third indoor heat exchanger 17 are arranged in the second indoor unit 3. As shown in FIG. 7, the third indoor heat exchanger 17 is disposed at a position where the same operation as that of the first indoor heat exchanger 10 is performed.

第３の室内熱交換器１７を通過した空気を低段側の第２室内熱交換器１３で冷却除湿することで、図８に示すように、冷却熱量の一部を高段側に移行することができる。図８は空気線図上に吸込みから吹出までの空気状態の変化を示している。破線は図１の構成による空気状態変化であり、状態Ａから状態Ｂに変化し、その単位空気質量あたりの熱量変化はΔｉ１で示される。また、この図７の構成による空気状態の変化は、状態Ａから第３室内熱交換器１７を通過して状態Ｃとなり、さらに第２室内熱交換器１３で冷却されて状態Ｂとなる。冷房時の熱量を送風機１５により熱交換器１７にて先ず処理し熱交換器１３における処理する顕熱負荷の熱量を下げることで効率が良い装置が得られる。   By cooling and dehumidifying the air that has passed through the third indoor heat exchanger 17 with the second indoor heat exchanger 13 on the lower stage side, a part of the cooling heat amount is shifted to the higher stage side as shown in FIG. be able to. FIG. 8 shows the change in air condition from suction to blowing on the air diagram. A broken line is an air state change according to the configuration of FIG. 1 and changes from the state A to the state B, and the heat amount change per unit air mass is indicated by Δi1. 7 changes from the state A through the third indoor heat exchanger 17 to the state C, and further cooled by the second indoor heat exchanger 13 to the state B. An apparatus with high efficiency can be obtained by first processing the amount of heat during cooling by the blower 15 in the heat exchanger 17 and lowering the amount of sensible heat load to be processed in the heat exchanger 13.

本発明の図１および図７の冷凍サイクルにおける室内ユニット３の入口空気、出口空気の状態はほぼ等しい。すなわち除湿量ΔＸは等しいが、低段蒸発圧力による処理熱量は、図８の方がΔｉ２分だけ小さくなることがわかる。図２で解説したように、ある一定の冷房負荷のもとで本実施の形態の冷凍サイクルＣＯＰは、トータル冷媒流量（Ｇ１＋Ｇ２）に占める低段冷媒流量Ｇ２の割合を小さくすることで向上するので、このような構成として高段側に必要冷却能力を移行することで本実施の形態の冷凍サイクルＣＯＰはさらに向上する。   The state of the inlet air and the outlet air of the indoor unit 3 in the refrigeration cycle of FIGS. 1 and 7 of the present invention is substantially equal. That is, although the dehumidification amount ΔX is equal, it can be seen that the amount of heat processed by the low stage evaporation pressure is smaller by Δi2 in FIG. As explained in FIG. 2, the refrigeration cycle COP of the present embodiment under a certain cooling load is improved by reducing the ratio of the low-stage refrigerant flow rate G2 to the total refrigerant flow rate (G1 + G2). In this configuration, the refrigeration cycle COP of the present embodiment is further improved by shifting the required cooling capacity to the higher stage side.

また、ビル用空調の場合強制的に外気を機械で取り入れることが法律などで定められている。このような場合、除湿負荷は室内空気よりも高湿度の外気を室内へ導入することにより発生する。強制換気により除湿負荷が発生している場合には、第２の室内ユニットに直接外気を処理させる、すなわち第２の室内ユニットに外気を導入させることで換気と同時に第２室内熱交換器１３の顕熱比を下げることができ、同一除湿量を得るために必要な低段側冷媒流量Ｇ２を小さくすることができる。すなわち高湿度の外気を除湿したほうが同一除湿量を得るために必要な低段側冷媒流量Ｇ２を小さくすることが出来る。図１１は外気を冷却除湿した場合と室内空気を冷却除湿した場合の違いを説明する図である。図１１にて状態Ａは室内空気状態であり、冷却除湿の過程を実線矢印で示している。この場合ΔＸの除湿量を得るためにはΔｉ１の冷却能力が必要である。一方状態Ｂは室内空気より高温高湿な外気で、冷却除湿の過程を破線矢印で示している。この場合ΔＸの除湿量を得るために必要な冷却能力はΔｉ２となる。図１１が示すようにΔｉ１とΔｉ２では高い湿度の空気を除湿するΔｉ２のほうが小さくなる。すなわち室内より高い湿度である外気を除湿すると同一除湿量を得るための低段冷媒流量Ｇ２は小さくてよい。強制換気による外気導入負荷が必然的に発生する場合は処理すべき冷房負荷に含まれるため、外気を直接潜熱熱交換器で処理して室内に導入することにより効率の良い装置が得られる。   In the case of air conditioning for buildings, the law stipulates that outside air is forcibly taken in by machines. In such a case, the dehumidifying load is generated by introducing outside air having higher humidity than room air into the room. When a dehumidifying load is generated by forced ventilation, the second indoor unit directly processes the outside air, that is, by introducing the outside air into the second indoor unit, the ventilation of the second indoor heat exchanger 13 is performed simultaneously with the ventilation. The sensible heat ratio can be lowered, and the low-stage refrigerant flow rate G2 necessary for obtaining the same dehumidification amount can be reduced. That is, the lower stage refrigerant flow rate G2 required to obtain the same dehumidifying amount can be reduced by dehumidifying the high humidity outside air. FIG. 11 is a diagram for explaining the difference between when the outside air is cooled and dehumidified and when the room air is cooled and dehumidified. In FIG. 11, the state A is an indoor air state, and the process of cooling and dehumidification is indicated by solid arrows. In this case, in order to obtain a dehumidification amount of ΔX, a cooling capacity of Δi1 is necessary. On the other hand, the state B is outside air having a higher temperature and humidity than room air, and the process of cooling and dehumidification is indicated by broken-line arrows. In this case, the cooling capacity necessary to obtain a dehumidification amount of ΔX is Δi2. As shown in FIG. 11, in Δi1 and Δi2, Δi2 for dehumidifying high-humidity air is smaller. That is, the low stage refrigerant flow rate G2 for obtaining the same dehumidifying amount when the outside air having a higher humidity than the room is dehumidified may be small. When the outside air introduction load due to forced ventilation is inevitably generated, it is included in the cooling load to be processed. Therefore, an efficient apparatus can be obtained by directly treating the outside air with the latent heat exchanger and introducing it into the room.

外気を室内に導入する構成として、図１および図７の第２の室内ユニット３をそのまま室内に配置して第２室内ユニット３に吸入する空気の入口を室外に設ける構造と、ユニット３全体を室外に配置する外調機としてユニットの空調された空気を室内へ吹出す吹出し口を室内側に設ける構造がある。外気を導入する入口を室外側に設ける構造が壁にあける穴は小さくて済むが、配管の工事などからは外調機構造が簡単である。また室内における発熱源の位置との関係など室内の空調性能によっても選択が左右される。   As a configuration for introducing the outside air into the room, the second indoor unit 3 shown in FIGS. 1 and 7 is arranged in the room as it is, and a structure in which an inlet for air sucked into the second indoor unit 3 is provided outside the room, There exists a structure which provides the blower outlet which blows off the air-conditioned air of a unit indoors as an external air conditioner arrange | positioned outdoors. The structure in which the inlet for introducing the outside air is provided outside the room requires a small hole in the wall, but the structure of the external air conditioner is simple from the construction work of the piping. The selection also depends on the indoor air conditioning performance such as the relationship with the position of the heat source in the room.

また図１２は潜熱熱交換器に対する予冷・再熱の構成説明図を示し、図１２の様に低段側であり第２の負荷側熱交換器である潜熱熱交換器１３に対し吸込み側の空気をあらかじめ別の顕熱熱交換器もしくは全熱熱交換器５６で予冷し、この潜熱熱交換器１３の吹出す空気で再熱させる構造にすると装置の効率が良い。図１２の様に全熱熱交換器５６を設け、図のような風路構造、即ち図示していない送風機および風路により室内空気を全熱熱交換器５６を通して予冷した後に潜熱熱交換器１３の吸込み空気とし、冷却除湿された温度の低い空気を全熱熱交換器５６のほかの風路を通して再熱する風路を採用する。このとき、潜熱熱交換器１３からの吹出空気は再度加熱されて室内に供給されるが、この温度上昇分は顕熱を処理する第１の室内熱交換器１０により冷却されるため、潜熱熱交換器１３の処理熱量は減少し、サイクルＣＯＰは向上する。図１２では例えば２７℃の室内空気を全熱熱交換器５６で２２℃まで予冷した後、低段蒸発器である潜熱熱交換器で１２℃まで冷却除湿する。２７℃から２２℃までの予冷はこの冷却除湿後の１２℃の空気でおこなわれる。一方１２℃の空気は２７℃の吸込み空気により再熱され１７℃で室内に吹出す。これにより強制外気取り入れ時の潜熱処理の効率を向上させることが出来る。   FIG. 12 is a diagram for explaining the configuration of pre-cooling / reheating for the latent heat exchanger. As shown in FIG. 12, the lower stage side and the second load side heat exchanger 13 are arranged on the suction side. If the air is precooled in advance with another sensible heat exchanger or the total heat exchanger 56 and reheated with the air blown out from the latent heat exchanger 13, the efficiency of the apparatus is good. As shown in FIG. 12, a total heat heat exchanger 56 is provided, and after the indoor air is pre-cooled through the total heat heat exchanger 56 by a wind path structure as shown in the figure, that is, a blower and a wind path not shown, the latent heat heat exchanger 13 An air path is employed in which the low-temperature air that has been cooled and dehumidified is reheated through another air path of the total heat exchanger 56. At this time, the air blown out from the latent heat exchanger 13 is heated again and supplied to the room, but this temperature rise is cooled by the first indoor heat exchanger 10 that processes the sensible heat. The amount of heat treated by the exchanger 13 is reduced, and the cycle COP is improved. In FIG. 12, for example, indoor air at 27 ° C. is pre-cooled to 22 ° C. by the total heat exchanger 56 and then cooled and dehumidified to 12 ° C. by a latent heat exchanger which is a low-stage evaporator. Precooling from 27 ° C. to 22 ° C. is performed with air at 12 ° C. after this cooling and dehumidification. On the other hand, the air at 12 ° C. is reheated by the intake air at 27 ° C. and blown into the room at 17 ° C. Thereby, the efficiency of the latent heat treatment at the time of forced outside air intake can be improved.

全熱熱交換器の代わりに第２の負荷熱交換器１３の吸込み側と吹き出し側にヒートパイプなどの顕熱熱交換器を設けても良い。図１３にこの状態を空気線図で示す。吸込み空気である状態Ａは予冷されて状態Ｃとなり冷却除湿される。除湿後の空気状態Ｂは再熱されて状態Ｄとなり再び室内へ吹出す。この予冷再熱が無い場合、Δｉ１に相当する冷却能力が必要なところ、予冷再熱することで除湿用熱交換器のＳＨＦが低くなり同一除湿量を得るための冷却能力はΔｉ３となりΔＩ２だけ冷却能力が不要になる。   Instead of the total heat heat exchanger, a sensible heat exchanger such as a heat pipe may be provided on the suction side and the outlet side of the second load heat exchanger 13. FIG. 13 shows this state with an air diagram. The state A that is the intake air is pre-cooled to state C and is cooled and dehumidified. The air state B after dehumidification is reheated to become the state D and blows out again into the room. If there is no precooling and reheating, a cooling capacity corresponding to Δi1 is required. However, by performing precooling and reheating, the SHF of the heat exchanger for dehumidification becomes low, and the cooling capacity for obtaining the same dehumidification amount becomes Δi3, which is cooled by ΔI2. The ability becomes unnecessary.

上述のように、高段蒸発圧力と低段蒸発圧力の２つの圧力で冷房負荷を処理する場合、極力低段側冷媒流量比を小さくすることでＣＯＰが向上するが、全冷房負荷に占める除湿負荷の割合がおおよそ０．４以上になると、その全冷房負荷のほとんどを低段蒸発圧力で処理しなければならなくなる。図３に示した顕熱比の特性からわかるように、蒸発温度をどんどん低下させても顕熱比はゼロにならず、通常の冷房設定条件である２５℃〜２８℃、相対湿度４０％〜６０％の範囲であれば０．６程度が下限値となる。このような除湿負荷の割合が比較的大きい冷房負荷では本発明の効果は小さく、全冷房負荷に対する除湿負荷の割合が０．２以下の空調対象に対して好適な空調システムということも出来る。例えば、第１の室内熱交換器と第２の室内熱交換器の合計冷房負荷を１００とし、そのうちの潜熱負荷を４０とする。ここで第２の室内熱交換器の顕熱比を０.６とすると第２の熱交換器で処理する負荷は、潜熱負荷４０／（１−ＳＨＦ）＝１００となり、第１の室内熱交換器で処理する冷房顕熱負荷が失われ冷房負荷の全てを第２の室内熱交換器で処理することになり省エネルギー効果は得られない。   As described above, when the cooling load is processed with two pressures of the high-stage evaporation pressure and the low-stage evaporation pressure, the COP is improved by reducing the low-stage side refrigerant flow ratio as much as possible, but the dehumidification occupies the total cooling load. When the load ratio is approximately 0.4 or more, most of the total cooling load must be processed at a low stage evaporation pressure. As can be seen from the characteristics of the sensible heat ratio shown in FIG. 3, the sensible heat ratio does not become zero even if the evaporation temperature is further decreased, and the normal cooling setting conditions of 25 ° C. to 28 ° C., relative humidity of 40% to In the range of 60%, about 0.6 is the lower limit. The effect of the present invention is small in such a cooling load having a relatively high dehumidifying load ratio, and it can be said that the air conditioning system is suitable for an air-conditioning target having a dehumidifying load ratio of 0.2 or less with respect to the total cooling load. For example, the total cooling load of the first indoor heat exchanger and the second indoor heat exchanger is set to 100, and the latent heat load thereof is set to 40. Here, when the sensible heat ratio of the second indoor heat exchanger is 0.6, the load to be processed by the second heat exchanger is latent heat load 40 / (1-SHF) = 100, and the first indoor heat exchange is performed. The cooling sensible heat load to be processed by the cooler is lost, and the entire cooling load is processed by the second indoor heat exchanger, so that the energy saving effect cannot be obtained.

通常の冷房条件、例えば室温２５−２８℃、湿度５０％程度であれば例えば露点温度以上の温度差で顕熱処理が出来る。一方露点温度以下では図８などの空気線図に示す様に通常の冷房条件である状態Ａから、飽和蒸気圧線の冷媒温度の方向へ傾いた方向の状態Ｂである状態変化線に下限があり０．６以下にはならない。即ち線熱負荷４０％以上は処理できない。このように潜熱と顕熱の負荷処理を４０対６０、すなわち２対３ではなく、それより小さな潜熱処理となるように各熱交換器面積比が設計されることが望ましい。これは空調機の省エネ効果が得られる負荷条件として、全冷房負荷に対する潜熱負荷比１０−２０パーセントを想定し、第２の室内熱交換器の顕熱比を０．５まで小さく出来たとすると、第２の室内熱交換器で処理する冷房負荷は全冷房負荷の２０−３０％となる。即ち第１の室内熱交換器に対する第２の熱交換器の熱処理能力が２０−３０％程度が望ましく、言い換えると顕熱処理側の熱交換器を潜熱処理側の熱交換器の３倍以上とすることである。このように顕熱熱交換器の熱処理能力を潜熱熱交換器の熱処理能力より大きくする際に、熱交換器そのものの伝熱面積比を３倍以上することと、あるいは負荷側送風手段の発生する風量、即ち送風ファンの回転数の両方で熱処理能力を３倍以上としても良い。更に第１の圧縮機と第２の圧縮機の最大容量を３倍以上に設定しておくと運転時にシステム全体の適合が取れて問題を発生することがない。このため第２の室内熱交換器で処理する冷房能力の割合が小さいこと、および図７のように処理空気の風路構成などにより第２の室内熱交換器の顕熱比を小さくすると良い。   Under normal cooling conditions, such as room temperature 25-28 ° C. and humidity of about 50%, sensible heat treatment can be performed with a temperature difference equal to or higher than the dew point temperature, for example. On the other hand, below the dew point temperature, there is a lower limit on the state change line which is the state B in the direction inclined from the normal cooling condition to the refrigerant temperature direction of the saturated vapor pressure line as shown in the air diagram of FIG. There is no less than 0.6. In other words, it is impossible to process a linear heat load of 40% or more. Thus, it is desirable to design each heat exchanger area ratio so that the latent heat and sensible heat load treatment is not 40:60, that is, 2: 3, but a smaller latent heat treatment. Assuming that the load condition for obtaining the energy saving effect of the air conditioner is 10 to 20% of the latent heat load ratio with respect to the total cooling load, and the sensible heat ratio of the second indoor heat exchanger can be reduced to 0.5, The cooling load to be processed by the second indoor heat exchanger is 20-30% of the total cooling load. That is, the heat treatment capacity of the second heat exchanger relative to the first indoor heat exchanger is preferably about 20-30%. In other words, the heat exchanger on the sensible heat treatment side is more than three times the heat exchanger on the latent heat treatment side. That is. As described above, when the heat treatment capacity of the sensible heat exchanger is made larger than the heat treatment capacity of the latent heat exchanger, the heat transfer area ratio of the heat exchanger itself is increased by three times or the load side blowing means is generated. It is good also considering heat processing capability as 3 times or more by both air volume, ie, the rotation speed of a ventilation fan. Furthermore, if the maximum capacities of the first compressor and the second compressor are set to be three times or more, the entire system can be adapted during operation without causing a problem. For this reason, it is preferable to reduce the sensible heat ratio of the second indoor heat exchanger due to the small ratio of the cooling capacity to be processed by the second indoor heat exchanger and the air path configuration of the processing air as shown in FIG.

以上の説明では冷媒としてハイドロフルオロカーボン、例えばＲ４１０Ａを使用する例を示したが、これに限定されることはない。例えば二酸化炭素のような自然冷媒を用いた場合においては、高段側蒸発圧力が臨界圧力に近いため、図２で説明したような高段側と低段側の蒸発エンタルピ差Δｈｅがほぼ等しいとはいえなくなる。このように蒸発温度が異なる場合、異なる蒸発圧力で蒸発器で利用できるエンタルピ差に差がある場合について次に説明する。   Although the example which uses hydrofluorocarbon, for example, R410A as a refrigerant | coolant was shown in the above description, it is not limited to this. For example, in the case where a natural refrigerant such as carbon dioxide is used, the high-stage evaporation pressure is close to the critical pressure, and therefore the high-stage and low-stage evaporation enthalpy differences Δhe described with reference to FIG. No more. Next, a case where there is a difference in enthalpy difference that can be used in the evaporator at different evaporation pressures when the evaporation temperatures are different will be described.

図９に本実施の形態の空気調和機の冷媒に二酸化炭素を用いた場合のＰ−ｈ線図を示す。図のように、臨界圧力付近では飽和ガス線が高圧側で大きく左の方へ湾曲するため、高段側で利用できるエンタルピ差が低段側より小さくなる（Δｈｅ１＜Δｈｅ２）。これは、全冷房負荷に対する低段側の処理熱量比に対して、全冷媒流量Ｇ１＋Ｇ２に占めるＧ２冷媒流量比がより小さくなることを示す。すなわち、必要な除湿量を得るために低段側能力比が全冷房能力の２０％必要とした場合、図２に示すようなＨＦＣ冷媒の物性ではΔｈｅが高段低段でほぼ等しいためＧ２／（Ｇ１＋Ｇ２）も２０％必要であるが、図９に示した二酸化炭素冷媒の物性によれば、Δｈｅ２＞Δｈｅ１であるため、Ｇ２比は２０％より小さくてよい。このことは、二酸化炭素冷媒の方がＨＦＣ冷媒よりＣＯＰ向上率が大きいことを示している。   FIG. 9 shows a Ph diagram when carbon dioxide is used as the refrigerant of the air conditioner of the present embodiment. As shown in the figure, near the critical pressure, the saturated gas line is bent largely to the left on the high pressure side, so that the enthalpy difference that can be used on the high stage side is smaller than that on the low stage side (Δhe1 <Δhe2). This indicates that the G2 refrigerant flow rate ratio in the total refrigerant flow rate G1 + G2 is smaller than the low-stage processing heat amount ratio with respect to the total cooling load. That is, when 20% of the total cooling capacity is required to obtain the necessary dehumidification amount, Δhe is almost equal between the high stage and the low stage in terms of the physical properties of the HFC refrigerant as shown in FIG. (G1 + G2) also needs 20%, but according to the physical properties of the carbon dioxide refrigerant shown in FIG. 9, since Δhe2> Δhe1, G2 ratio may be smaller than 20%. This indicates that the carbon dioxide refrigerant has a higher COP improvement rate than the HFC refrigerant.

さらに、前述のように、低段側冷媒流量Ｇ２が全冷媒流量の２０％以下であることがこの実施の形態の空気調和機に好適な条件であるので、図１０に示すような冷媒回路とし、高圧から中圧までの膨張動力を回収することで低段側圧縮動力をゼロとすることができる。以下に図１０を参照して詳細に説明する。また、図２と同様の部分については同一の番号を付し、説明を省略する。特に超臨界状態の存在する冷媒、例えば二酸化炭素のような場合圧力が高く圧力差が存在するため膨張動力の回収が有効である。   Furthermore, as described above, the low-stage refrigerant flow rate G2 is 20% or less of the total refrigerant flow rate, which is a suitable condition for the air conditioner of this embodiment. By collecting the expansion power from high pressure to medium pressure, the low-stage compression power can be made zero. This will be described in detail with reference to FIG. Moreover, the same number is attached | subjected about the part similar to FIG. 2, and description is abbreviate | omitted. In particular, in the case of a refrigerant in a supercritical state, such as carbon dioxide, recovery of expansion power is effective because the pressure is high and there is a pressure difference.

第２室内ユニット３には、膨張動力回収手段２１を内蔵している。膨張動力回収手段２１は膨張部２２と圧縮部２３を備え、同軸で連結されている。室外ユニット１より流入する高圧液冷媒を中圧まで減圧する際の膨張動力を回収して圧縮部２３で圧縮仕事を行う。この図１０の例では、中圧まで減圧した冷媒をさらに第２減圧手段１２により減圧し、第２室内熱交換器１３で室内空気を冷却除湿した後、圧縮部２３で再び中圧まで昇圧している。 The second indoor unit 3 incorporates expansion power recovery means 21. The expansion power recovery means 21 includes an expansion portion 22 and a compression portion 23, and is connected coaxially. The expansion power when the high-pressure liquid refrigerant flowing from the outdoor unit 1 is reduced to an intermediate pressure is recovered, and the compression unit 23 performs the compression work. In the example of FIG. 10, the refrigerant depressurized to the medium pressure is further depressurized by the second depressurization means 12, the indoor air is cooled and dehumidified by the second indoor heat exchanger 13, and then the pressure is increased again to the medium pressure by the compression unit 23. ing.

ここで圧縮される冷媒流量は前述のとおり、全冷媒流量の２０％〜３０％程度であるので、二酸化炭素の物性および膨張動力回収効率からみて十分昇圧できる流量であり、このような構成とすることで低段側は膨張動力だけとなり他の入力をゼロにすることができる。   Since the refrigerant flow rate compressed here is about 20% to 30% of the total refrigerant flow rate as described above, the flow rate can be sufficiently increased in view of the physical properties of carbon dioxide and the expansion power recovery efficiency. In this way, the low stage side has only expansion power, and other inputs can be made zero.

これらの構成における冷凍サイクルＣＯＰ向上方法は、空気調和に限定されるものではない。冷凍と冷蔵で異蒸発温度を発生させるシステムにおいても全く同じ原理でＣＯＰを向上することができる。   The refrigeration cycle COP improvement method in these configurations is not limited to air conditioning. Even in a system that generates different evaporation temperatures by freezing and refrigeration, the COP can be improved by the same principle.

本発明のように熱負荷を顕熱処理と潜熱処理に分けた上で処理した上に総合的に送風機制御や各処理ユニットの配置や潜熱処理の熱交換器を顕熱処理の熱交換器より大幅に小さくするなどの構造の組み合わせ他で総合的な効率化を図ることによりその効率向上が著しく拡大できる。これにより地球環境保護対策として有効な冷凍サイクルが得られ、空気調和機、冷蔵庫、冷凍倉庫、コンビニやスーパーストアにおける冷凍冷蔵システムなどに有効である。更に加えて２酸化炭素（ＣＯ２）のような自然冷媒を使用する冷凍サイクルにおいては圧力が高く図９のように凝縮時に超臨界状態となり特性が異なるとともに冷媒を循環させる配管や容器の耐圧体力に配慮が必要であるが地球環境上Ｒ２２等のフロンのようにオゾン層を破壊することが無く、代替フロンのように地球温暖化への影響も少ない。したがって炭酸ガス冷媒で効率向上が求められ図１０のように高い圧力に基づく大きな圧力差を膨張動力として利用することが望ましい。次にこのような自然冷媒を使用した冷凍サイクルにおけるほかの高効率対策を説明する。   The heat load is divided into sensible heat treatment and latent heat treatment as in the present invention, and the overall control of the blower, the arrangement of each processing unit, and the heat exchanger of the latent heat treatment are significantly larger than the heat exchanger of the sensible heat treatment. The efficiency improvement can be remarkably expanded by improving the overall efficiency by combining the structures such as reducing the size. As a result, a refrigeration cycle effective as a global environmental protection measure is obtained, which is effective for air conditioners, refrigerators, refrigeration warehouses, refrigeration systems in convenience stores and superstores, and the like. In addition, in a refrigeration cycle that uses a natural refrigerant such as carbon dioxide (CO2), the pressure is high, and as shown in FIG. Although consideration is necessary, the ozone layer is not destroyed like chlorofluorocarbons such as R22 on the global environment, and it has little impact on global warming like alternative chlorofluorocarbons. Therefore, the efficiency improvement is required with the carbon dioxide refrigerant, and it is desirable to use a large pressure difference based on the high pressure as the expansion power as shown in FIG. Next, another high efficiency measure in the refrigeration cycle using such a natural refrigerant will be described.

大きな圧力差を利用して効率向上を図る機構として膨張動力回収手段２１、例えば膨張部２２と圧縮部２３を備え、同軸で連結され同一回転数で回転する機械的な動力回収構造がある。このような膨張動力回収手段２１は従来絞り装置で摩擦損失や渦損失の形で無駄に熱として捨てられていた流体の膨張動力体積膨張に伴う膨張仕事である膨張機出入り口でのエンタルピー差の断熱熱落差を回収して機械エネルギー例えば回転動力や往復動力に変換し圧縮機の圧縮仕事に活用する。膨張動力回収手段２１の膨張部２２で駆動する圧縮部２３は第２の圧縮機として室内の空調に使用する。冷房時には潜熱処理を行う第２の負荷側熱交換器への冷媒を循環させ、暖房時には高圧側の配管に切換えられた膨張部２２にて動力を回収し低圧側の配管へ冷媒を循環させて第２の冷媒サイクルの凝縮熱を室内へ吹出させる負荷処理を行い効率の向上に役立てる。このように高圧で超臨界となる冷媒では高圧側の体積膨張が大きくなると共に、即ち密度減少が大きくなり低圧の渇き度が大きくなるため膨張損失が大きいので膨張動力を利用することが省エネルギーに有効である。その上この膨張動力のみを第２圧縮機の駆動力である膨張部２２に利用し再び冷媒を高圧側に戻すことで更に省エネルギーを図ることが出来る。以上の様に冷媒の物性である圧力差や密度などより得られる運動エネルギー、即ち流体の堆積流量もしくは流速を利用する駆動機構であれば度のような構造でも使用できる。   As a mechanism for improving efficiency by using a large pressure difference, there is an expansion power recovery means 21, for example, an expansion portion 22 and a compression portion 23, and there is a mechanical power recovery structure that is connected coaxially and rotates at the same rotational speed. Such an expansion power recovery means 21 is an insulation of the enthalpy difference at the entrance of the expander, which is the expansion work associated with the expansion power volume expansion of the fluid that has been wasted as heat in the form of friction loss and vortex loss in the conventional throttle device. The heat drop is recovered and converted into mechanical energy such as rotational power or reciprocating power and used for the compression work of the compressor. The compression part 23 driven by the expansion part 22 of the expansion power recovery means 21 is used for indoor air conditioning as a second compressor. During cooling, the refrigerant is circulated to the second load-side heat exchanger that performs latent heat treatment, and during heating, power is recovered by the expansion unit 22 switched to the high-pressure side piping, and the refrigerant is circulated to the low-pressure side piping. A load process for blowing the heat of condensation of the second refrigerant cycle into the room is performed to help improve efficiency. In such a supercritical refrigerant at high pressure, the volume expansion on the high pressure side increases, that is, the density loss increases and the degree of thirst at low pressure increases, so the expansion loss is large, so the use of expansion power is effective for energy saving. It is. In addition, energy can be further saved by using only this expansion power for the expansion section 22 which is the driving force of the second compressor and returning the refrigerant to the high pressure side again. As described above, any structure can be used as long as it is a drive mechanism that uses the kinetic energy obtained from the pressure difference or density, which is the physical property of the refrigerant, that is, the fluid deposition flow rate or flow velocity.

以上のように本発明の図１０の構成では膨張部２２にて高圧の流体が低圧に膨張する際に膨張仕事を行い、これに伴い従来熱として消費していた仕事を圧縮部２３の駆動力としている。特に高圧で超臨界状態となる炭酸ガスのような冷媒は、物性上比較的大きな回収動力が得られる。なおこのような仕事を行う自然冷媒として空気、窒素、ヘリウムや、あるいは炭酸ガスなどの自然冷媒とＨＦＣ他の冷媒との混合冷媒でも、異なる蒸発圧力で蒸発器に利用できるエンタルピ差に差がある冷媒であれば同様に効果があることは言うまでもない。一例としての膨張動力回収手段２１は膨張部２２と圧縮部２３が密閉容器の内部にそれぞれのスクロール構造が直結され一体となった構造で膨張動力をそのまま直接圧縮する様に形成された簡単な構造である。即ち、密閉容器中の軸方向両端に膨張用固定スクロールと第２圧縮用固定スクロールが対抗されて固定配置され、容器中心には回転軸を支持する軸受が設けられている。容器中央部に各固定スクロールと噛合い揺動回転する膨張用および圧縮用の揺動スクロールが回転軸により駆動され一体の回転子として形成されている。即ち膨張部での膨張動力で駆動され揺動回転し圧縮部で圧縮仕事をしているだけの構造である。このため高圧管に接続され密閉容器に設けられた膨張吸入管から膨張用固定スクロールを通して膨張用揺動スクロールとの噛合い空間の軸中心に近い個所に高圧冷媒が吸入されこの揺動スクロールを揺動させながら回転子を回転させ容器壁部に近い空間から膨張吐出管へ吐出され再び高圧配管に戻される。次に膨張機１２を通過後第２の負荷側熱交換器にて蒸発され圧力の低下した冷媒は圧縮部２３への第２圧縮吸入管から圧縮用固定スクロールを介して圧縮用揺動スクロールとの噛合い空間の軸中心より外れた外周側にこのガス冷媒が吸入され圧縮部２３のスクロールが回転軸を介して駆動揺動されるにともない軸中心に近い内側の噛合い空間からガス圧力が上昇されて吐出管から低圧配管に戻される。すなわち密閉容器内のそれぞれのスクロール構造へ冷媒を出し入れをするだけで膨張動力を圧縮仕事に変換でき無駄のない動作が簡単な構造で行える。なおここではスクロール圧縮機を取り上げて説明したが往復駆動型圧縮機であっても良い。更にこのような容積型の駆動装置でなく流体の体積流量や流速を利用した羽根車形状の駆動装置にすると容積型で問題になる余剰分対策のバイパス弁などが不要になる。   As described above, in the configuration of FIG. 10 according to the present invention, expansion work is performed when a high-pressure fluid expands to low pressure in the expansion section 22, and the work conventionally consumed as heat is driven by the driving force of the compression section 23. It is said. In particular, a refrigerant such as carbon dioxide gas that becomes a supercritical state at a high pressure can provide a relatively large recovery power in terms of physical properties. In addition, as a natural refrigerant for performing such work, there is a difference in enthalpy difference that can be used for an evaporator at different evaporating pressures even in a mixed refrigerant of natural refrigerant such as air, nitrogen, helium, or carbon dioxide gas and HFC or other refrigerants. Needless to say, a refrigerant is also effective. The expansion power recovery means 21 as an example is a simple structure in which the expansion portion 22 and the compression portion 23 are directly connected to the inside of the sealed container so as to directly compress the expansion power as it is. It is. That is, the expansion fixed scroll and the second compression fixed scroll are fixedly arranged opposite to each other in the axial direction in the sealed container, and a bearing for supporting the rotating shaft is provided at the center of the container. In the center of the container, a swing scroll for expansion and compression that meshes with each fixed scroll and rotates is driven by a rotary shaft and formed as an integral rotor. That is, it is a structure that is driven by the expansion power in the expansion portion and swings and rotates and performs compression work in the compression portion. For this reason, high-pressure refrigerant is sucked from the expansion suction pipe connected to the high-pressure pipe through the fixed expansion scroll through the fixed fixed scroll and into the position near the axial center of the meshing space with the expansion swing scroll, and the swing scroll is shaken. The rotor is rotated while being moved, and is discharged from the space close to the container wall to the expansion discharge pipe and returned to the high-pressure pipe again. Next, after passing through the expander 12, the refrigerant whose pressure has been reduced by evaporation in the second load side heat exchanger passes through the second compression suction pipe to the compression unit 23 through the compression fixed scroll and the compression swing scroll. As the gas refrigerant is sucked into the outer peripheral side deviating from the shaft center of the meshing space and the scroll of the compression portion 23 is driven and oscillated via the rotating shaft, the gas pressure is generated from the meshing space inside the shaft center. It is raised and returned from the discharge pipe to the low pressure pipe. That is, the expansion power can be converted into compression work by simply putting the refrigerant in and out of the respective scroll structures in the sealed container, and the operation without waste can be performed with a simple structure. Although the scroll compressor has been described here, a reciprocating compressor may be used. Further, if the impeller-shaped drive device using the volumetric flow rate and flow velocity of the fluid is used instead of the positive displacement type drive device, a bypass valve or the like for preventing the surplus which becomes a problem with the positive displacement type becomes unnecessary.

圧力差利用の動力回収としてこのような機械的な膨張動力回収手段では無く、エジェクタを使用することも出来る。図１４にこのエジェクタの原理説明図を示す。図１４に示すようにノズル３１から圧力の高い冷媒を減圧させて噴出させる時にノズル周囲から圧力の低い冷媒を吸引して混合し中圧の冷媒とするものである。図の駆動流は凝縮された高圧液冷媒であって、この駆動流が減圧されながら高速ニ相噴流となってノズル３１から噴出す際に最も圧力が低い状態となり吸引流を引き込む。合流した駆動流と吸引流は混合部３２で均質に混合されディフューザ部３３にて減速し圧力を回復してから下流へと進行する。エジェクタ３０は通常の減圧手段、例えば膨張弁等では冷媒の流れが弁などとの摩擦によりエネルギーを失うのに対し、このエネルギーを吸引流の昇圧に利用するので、駆動流を減圧した際にエンタルピが減少し蒸発器で利用できるエンタルピが拡大できるとともに、圧縮機吸入圧力よりも低い冷媒流れを生成できる。又このエジェクタのノズル３１部分にニードル弁を設けることにより第３室内熱交換器１７への冷媒流量を制御することが出来る。即ちエジェクタはニードル弁を設けることで単に顕熱熱交換器１７の膨張弁の役割を果たすことになる。   Instead of such mechanical expansion power recovery means, it is possible to use an ejector for recovering power using pressure difference. FIG. 14 is a diagram for explaining the principle of this ejector. As shown in FIG. 14, when the high pressure refrigerant is decompressed and ejected from the nozzle 31, the low pressure refrigerant is sucked and mixed from the periphery of the nozzle to obtain a medium pressure refrigerant. The driving flow shown in the figure is a condensed high-pressure liquid refrigerant. When the driving flow is reduced in pressure and becomes a high-speed two-phase jet, the pressure becomes the lowest when it is ejected from the nozzle 31, and the suction flow is drawn. The combined driving flow and suction flow are mixed uniformly in the mixing section 32, decelerated in the diffuser section 33, recovered in pressure, and then proceed downstream. The ejector 30 loses energy due to the friction of the refrigerant flow with a normal pressure reducing means, for example, an expansion valve, etc., whereas this energy is used for boosting the suction flow. Therefore, when the drive flow is decompressed, the enthalpy As a result, the enthalpy that can be used in the evaporator can be expanded and a refrigerant flow lower than the compressor suction pressure can be generated. Further, by providing a needle valve at the nozzle 31 portion of the ejector, the refrigerant flow rate to the third indoor heat exchanger 17 can be controlled. That is, the ejector simply serves as an expansion valve of the sensible heat exchanger 17 by providing the needle valve.

図１５にエジェクタ３０を使用した冷凍サイクルの冷房時の構成図を示す。図の冷凍サイクルは圧縮機１４から吐出され凝縮器である室外熱交換器５にて凝縮された高圧冷媒が室外ユニット１から送り出され第２の室内ユニット３にてその一部がエジェクタ３０のノズル部３１で減圧され蒸発器である第３室内熱交換器１７にて蒸発し室内熱負荷の主として顕熱を処理する。高圧冷媒の残りは第３減圧手段１２で減圧され第２室内熱交換器１３にて蒸発し主として潜熱の処理を行う。第２室内熱交換器から出た低温低圧の冷媒はエジェクタ３０のノズル３１周辺から高圧冷媒に吸引され中間圧冷媒となり第３室内熱交換器１７にて蒸発し圧縮機４に吸入される。このようにエジェクタ３０はその吸引部が接続された第２の室内熱交換器１３の蒸発圧力を圧縮機４の吸入圧力よりも低くなるように動作させている。これにより、第３の熱交換器１７において蒸発温度が露点温度以上である場合でも、第２室内熱交換器の蒸発温度が室内空気の露点温度以下となり除湿することが可能になる。   FIG. 15 shows a configuration diagram during cooling of a refrigeration cycle using the ejector 30. In the refrigeration cycle shown in the figure, the high-pressure refrigerant discharged from the compressor 14 and condensed in the outdoor heat exchanger 5 that is a condenser is sent out from the outdoor unit 1, and a part of the high-pressure refrigerant is ejected from the ejector 30 in the second indoor unit 3. The pressure is reduced in the section 31 and evaporated in the third indoor heat exchanger 17 as an evaporator, and mainly sensible heat of the indoor heat load is processed. The remainder of the high-pressure refrigerant is depressurized by the third depressurizing means 12, evaporated in the second indoor heat exchanger 13, and mainly processed for latent heat. The low-temperature and low-pressure refrigerant discharged from the second indoor heat exchanger is sucked into the high-pressure refrigerant from the vicinity of the nozzle 31 of the ejector 30 and becomes intermediate-pressure refrigerant, which is evaporated in the third indoor heat exchanger 17 and sucked into the compressor 4. Thus, the ejector 30 is operated so that the evaporation pressure of the second indoor heat exchanger 13 to which the suction part is connected is lower than the suction pressure of the compressor 4. As a result, even when the evaporation temperature in the third heat exchanger 17 is equal to or higher than the dew point temperature, the evaporation temperature of the second indoor heat exchanger becomes equal to or lower than the dew point temperature of the room air, and dehumidification is possible.

次に図1６にエジェクタを使用した冷凍サイクルのＰｈ線図を、図１７はこの動作を説明する冷凍サイクル構成図を示す。図１６は模擬的な動作を示すものであって、凝縮器で凝縮した高圧液冷媒（状態Ａ）は、その一部が第２の減圧手段１２により状態Ｂまで減圧され第２の室内熱交換器１３にて室内空気を冷却除湿しながら蒸発して状態Ｃとなる。この状態Ｃの低温ガスの冷媒はエジェクタ３０の吸引流となって吸引され混合部３２に引き込まれる。一方状態Ａの高圧液冷媒はエジェクタ３０の駆動流として流入し吸引流を吸引する昇圧仕事、即ち図１６の状態Ｃから状態Ｄへの昇圧によりエンタルピを減少させながら減圧されて状態Ｅとなる。エジェクタ内では状態Ｄと状態Ｅの冷媒が合流し状態Ｆとなって第３室内熱交換器１７へ流入し室内空気を冷却しながら蒸発して状態Ｇとなる。但しこの動作は説明の都合状、状態Ｄや状態Ｅを仮定したもので、実際には状態ＣからＦへ、状態ＡからＦへ特性的には動いていく。エジェクタ内では駆動流が減圧されながら高速ニ相噴流となってノズル３１から噴出す際に最も圧力が低い状態となり吸引流を引き込むのでエジェクタの外の圧力は昇圧された状態Ｆで蒸発しガス配管８を介して第１の圧縮機４に吸引される。   Next, FIG. 16 shows a Ph diagram of a refrigeration cycle using an ejector, and FIG. 17 shows a refrigeration cycle configuration diagram for explaining this operation. FIG. 16 shows a simulated operation. A part of the high-pressure liquid refrigerant condensed in the condenser (state A) is decompressed to the state B by the second decompression means 12 and is subjected to the second indoor heat exchange. In room 13, the room air is evaporated while being cooled and dehumidified, so that state C is obtained. The refrigerant of the low-temperature gas in the state C is sucked into the ejector 30 and drawn into the mixing unit 32. On the other hand, the high-pressure liquid refrigerant in the state A flows into the drive flow of the ejector 30 and is depressurized while reducing the enthalpy by the pressurization work for sucking the suction flow, that is, the pressure increase from the state C to the state D in FIG. In the ejector, the refrigerants in the state D and the state E are merged to be in the state F, flow into the third indoor heat exchanger 17 and evaporate while cooling the indoor air to be in the state G. However, this operation assumes the state D and the state E for convenience of explanation, and actually moves characteristically from the state C to F and from the state A to F. In the ejector, the drive flow is reduced in pressure while the high-speed two-phase jet is ejected from the nozzle 31, the pressure is lowest and the suction flow is drawn. Therefore, the pressure outside the ejector evaporates in the pressurized state F, and the gas pipe The air is sucked into the first compressor 4 through 8.

図１７では第２室内送風機１５にて顕熱を主体に処理する第３室内熱交換器１７と潜熱を主体に処理する第２室内熱交換器１３を直列送風で室内空気との熱交換を行う構造を説明しているがこれは潜熱熱交換器の効率特性向上のためであって、この室内熱交換器への送風を熱交換面積を考えて１台の送風機で並列に送風しても良いし、各個別の熱交換器それぞれに独立した送風機を設けてそれぞれ運転状況により独立した制御を行うことでも良い。図１７とは異なる構成で１台の室内ユニット中に潜熱熱交換器と顕熱熱交換器を収納してエジェクタを使用する構成を図１８に、又そのＰｈ線図を図１９に示す。第１圧縮機である主圧縮機４から吐出され室外に設けられた熱源側熱交換器である凝縮器５で凝縮された高圧液冷媒、状態Ａは先ずエジェクタで減圧され状態Ｂとなる。この状態Ｂの温度は室内露点温度以上である。この状態Ｂの冷媒の一部が分岐され更に膨張弁７１で室内露点温度以下である状態Ｃ迄減圧する。第２室内熱交換器１３にて蒸発して主として潜熱を処理した冷媒は状態Ｄとなりエジェクタにより昇圧されて再び露点温度以上の状態Ｂの温度となる。この状態Ｂの冷媒は第３室内熱交換器１７で顕熱処理主体の蒸発を行い状態Ｅとなって圧縮機４へ吸引される。   In FIG. 17, the third indoor heat exchanger 17 that mainly processes sensible heat in the second indoor blower 15 and the second indoor heat exchanger 13 that mainly processes latent heat exchange heat with room air by serial air blowing. Although the structure is described, this is to improve the efficiency characteristics of the latent heat exchanger, and the air to the indoor heat exchanger may be blown in parallel by one blower in consideration of the heat exchange area. In addition, an independent blower may be provided for each individual heat exchanger, and independent control may be performed depending on the operation status. FIG. 18 shows a configuration in which a latent heat exchanger and a sensible heat exchanger are accommodated in one indoor unit with a configuration different from that in FIG. 17, and a Ph diagram thereof is shown in FIG. The high-pressure liquid refrigerant discharged from the main compressor 4 that is the first compressor and condensed in the condenser 5 that is the heat source side heat exchanger provided outside the chamber, the state A is first reduced to the state B by the ejector. The temperature in state B is equal to or higher than the indoor dew point temperature. A part of the refrigerant in this state B is branched and further decompressed by the expansion valve 71 to a state C that is not more than the indoor dew point temperature. The refrigerant that has evaporated in the second indoor heat exchanger 13 and mainly processed the latent heat is in the state D and is pressurized by the ejector, and again reaches the temperature of the state B that is equal to or higher than the dew point temperature. The refrigerant in the state B is evaporated in the sensible heat treatment main body in the third indoor heat exchanger 17 to become the state E and is sucked into the compressor 4.

図１９の構成では膨張弁７１の入口側冷媒、即ち状態Ｂの冷媒が二相であるため膨張弁通過時の冷媒の流による振動や騒音が問題になるような場合は、実用的な構造として図２０のような構成とすると良い。図２０は図１９の冷凍サイクルの一部を説明する図であって、エジェクタの下流側に分配器７２を設け、第２の室内熱交換器側には絞り量が大きなキャピラリチューブ、即ち細めのキャピラリを接続させて室内露点温度以下まで減圧してから第２の室内熱交換器にて主として潜熱負荷の処理を行いエジェクタ３０の吸引流とする。第３の室内熱交換器１７への接続は絞り量の小さな太目のキャピラリチューブとして室内露点温度以上の冷媒温度を確保して主として顕熱負荷の処理を行う。この構成により簡単な構造でエジェクタを利用して効率向上が出来る冷媒サイクルシステムを形成させることができる。   In the configuration of FIG. 19, since the refrigerant on the inlet side of the expansion valve 71, that is, the refrigerant in the state B is two-phase, if vibration or noise due to the refrigerant flow when passing through the expansion valve becomes a problem, a practical structure is used. A configuration as shown in FIG. 20 is preferable. FIG. 20 is a diagram for explaining a part of the refrigeration cycle of FIG. 19, in which a distributor 72 is provided on the downstream side of the ejector, and a capillary tube having a large throttle amount, that is, a narrower tube, is provided on the second indoor heat exchanger side. After the capillary is connected and the pressure is reduced to the room dew point temperature or lower, the latent heat load is mainly processed in the second indoor heat exchanger to obtain the suction flow of the ejector 30. The connection to the third indoor heat exchanger 17 is a thick capillary tube with a small throttle amount, which ensures a refrigerant temperature equal to or higher than the indoor dew point temperature and mainly performs sensible heat processing. With this configuration, it is possible to form a refrigerant cycle system that can improve efficiency by using an ejector with a simple structure.

なお図１８、図２０では室内ユニット３４に収納する第２の室内熱交換器１３と第３の室内熱交換器１７とを直列に室内空気を流す送風機１５を設けた構造を説明しているが、個々の熱交換器それぞれ独立に送風機を設けても良いし、１つの送風機で並列に送風させても良いことは既に説明した通りである。この後者の室内ユニット構造図を図２１に示す。図２１の空気調和機は室内の天井５０に埋め込む構造で、天井埋め込みカセット５１の内部に室内の空気を中心から吸込む送風機５２を収納し回転して周囲に吹出す空気を熱交換器で冷却して再び室内へ吹出し流として吹出している。熱交換器１０、１３は送風機５２の周囲を覆う形で設けられ下部に設けられたドレンパン５４にて結露水を集めている。室内に面した前面側の吸気流取り入れ口にはフィルタ５３が設けられ、又カセット５１の背部、即ち天井裏に隠れた位置にエジェクタ３０や暖房時にエジェクタ吸入路やバイパス回路を開閉する開閉弁などを収納して固定し内部を防音兼用の断熱材に覆われた収納箱が設けられている。上部熱交換器１０（１７）は主として顕熱を主体に処理しており。熱交換面積の大半を占めている。この熱交換器に流す冷媒の温度は露点温度以上であり露付きがほとんどないためフィンに水を流す考慮は不要で、このためフィンを上下方向に並べて配置する必要もない。最も効率の良い熱交換器構造を採用すれば良く、例えば冷媒を通す伝熱管として上下にヘシャゲさせて水平方向に広げた通風方向に複数の扁平管を設け、その上下の伝熱管の間にコルゲートフィンを設け効率向上を図っても良い。即ち扁平管の間にコルゲート状フィンを挟みこみ熱伝達させる構造である。一方下部熱交換器１３は潜熱を処理するため冷媒の温度を露点温度以下にしており結露水の処理を行いやすいプレートフィンを上下方向に配列させる構造などが採用され、しかもドレンパン５４に近接して設けている。   18 and 20, the structure in which the blower 15 that flows the indoor air in series between the second indoor heat exchanger 13 and the third indoor heat exchanger 17 housed in the indoor unit 34 is described. As described above, each of the heat exchangers may be provided with a blower independently, or may be blown in parallel with a single blower. FIG. 21 shows the structure of the latter indoor unit. The air conditioner of FIG. 21 has a structure embedded in the ceiling 50 of the room. A blower 52 for sucking indoor air from the center is housed in the ceiling-embedded cassette 51, and the air blown around is cooled by a heat exchanger. Then it is blown out into the room again. The heat exchangers 10 and 13 are provided so as to cover the periphery of the blower 52 and collect condensed water by a drain pan 54 provided in the lower part. A filter 53 is provided in the intake air intake on the front side facing the room, and the back of the cassette 51, that is, an opening / closing valve that opens and closes the ejector suction path and bypass circuit during heating in a position hidden behind the ceiling, etc. A storage box is provided, which is housed and fixed and covered with a heat insulating and heat insulating material. The upper heat exchanger 10 (17) mainly processes sensible heat. It occupies most of the heat exchange area. Since the temperature of the refrigerant flowing through the heat exchanger is equal to or higher than the dew point temperature and there is almost no dew, there is no need to consider flowing water through the fins. Therefore, it is not necessary to arrange the fins side by side. The most efficient heat exchanger structure may be adopted. For example, a plurality of flat tubes are installed in the air flow direction that is squeezed up and down and spread horizontally as a heat transfer tube through which the refrigerant passes, and between the upper and lower heat transfer tubes Fins may be provided to improve efficiency. That is, the corrugated fin is sandwiched between the flat tubes to transfer heat. On the other hand, the lower heat exchanger 13 employs a structure in which the temperature of the refrigerant is set to the dew point temperature or less in order to process latent heat, and plate fins that are easy to treat condensed water are arranged in the vertical direction. Provided.

又下部熱交換器１３は主として潜熱処理を行うため伝熱管の分流数を増やす多岐分流構造としても良い。熱交換器における効率向上には配管径を小さくすればするほどベターであり、したがって小面積でも効率向上が出来、分流数が増えることになるが潜熱処理主体の熱交換器ではこの分流数を増やしてもアンバランスが増大する恐れもほとんどない。このような細管熱交換器に対してはワイヤー状のフィンを伝熱細管に巻きつけたり接触させたりして、且つこのワイヤーを斜めに張り巡らせて水が流れる構造にすると良い。この様に顕熱処理の熱交換器と潜熱処理の熱交換器はその熱交換処理の面積である寸法が大きく異なるだけでなく、それぞれの特徴や効率向上を目的にした伝熱管の寸法や形状、フィンの寸法や形状など全く異なるものの組合せにすることが出来、またこの組合せ接続部に特別な配慮は不要で図２１の様に単純に上下に配置するようなもので良い。   Moreover, since the lower heat exchanger 13 mainly performs a latent heat treatment, it may have a multi-divided flow structure that increases the number of diverted heat transfer tubes. In order to improve efficiency in heat exchangers, the smaller the pipe diameter, the better. Therefore, efficiency can be improved even in a small area, and the number of diverts will increase. However, there is little fear of increasing unbalance. For such a thin tube heat exchanger, it is preferable that a wire-like fin is wound around or brought into contact with the heat transfer thin tube, and the wire is stretched obliquely so that water flows. In this way, the heat exchanger for sensible heat treatment and the heat exchanger for latent heat treatment are not only greatly different in size, which is the area of the heat exchange treatment, but also the size and shape of the heat transfer tubes for the purpose of improving their characteristics and efficiency, A combination of completely different fin dimensions and shapes can be used, and no special consideration is required for the combination connecting portion, and the fins can be simply arranged vertically as shown in FIG.

以上図１などの冷媒サイクルの構成として冷房主体で説明してきたが、冷媒の流を変更する四方弁等を設けることにより室内の空気調和を暖房にすることは容易である。この構成の一例を図２２に示す。図２２は室外ユニット１に熱源側サイクルである高段圧縮機４と室外熱交換器５を配置し、第１の四方弁３５にて冷媒の流を反対方向に切換えられるようにしている。第１の室内ユニット２は顕熱処理主体の第１熱交換器１０を複数設けた構造とし、第２の室内ユニットは潜熱処理主体の第２室内熱交換器１３を設けている。それぞれの熱交換器には減圧手段９、１２を直列に配置すると共に、第２の室内ユニット３には低段圧縮機１４およびこの吐出方向を変更できる第２の四方弁３６を設けている。これにより室外熱交換器を凝縮器にも蒸発器にも使用でき、したがって室内に配置した熱交換器を蒸発器として冷房に、凝縮器として暖房にも利用できる。また図１５の構成図の様にエジェクタを使用するサイクルの場合でもガス配管８を第２の室内熱交換器１３まで延長して反膨張弁側に接続しこの延長配管に第１の開閉弁を設けると共にこの第２の室内熱交換器からエジェクタ３０への吸引流の配管にも第２の開閉弁を設け熱源側の四方弁３５の切換に合わせて開閉弁を開閉させるなどの構成で冷房、暖房とも可能である。即ち室内を冷房空調したい場合は第１の開閉弁を閉じて第２の開閉弁を開放させることより既に説明済みの動作が行われる。一方熱源側を逆転させた場合は第１の開閉弁を開放し、第２の開平弁を閉鎖させると、冷媒は熱交換器１３、１７に並列に供給され両方の熱交換器とも凝縮器として動作する。なおエジェクタ３０は何の動作にも関係しない冷媒を通過させる配管となる。   As described above, the configuration of the refrigerant cycle in FIG. 1 and the like has been described mainly with respect to cooling. However, it is easy to heat indoor air conditioning by providing a four-way valve or the like that changes the flow of the refrigerant. An example of this configuration is shown in FIG. In FIG. 22, the high-stage compressor 4 and the outdoor heat exchanger 5, which are heat source side cycles, are arranged in the outdoor unit 1, and the refrigerant flow can be switched in the opposite direction by the first four-way valve 35. The first indoor unit 2 has a structure in which a plurality of first heat exchangers 10 mainly composed of sensible heat treatment are provided, and the second indoor unit is provided with a second indoor heat exchanger 13 mainly composed of latent heat treatments. In each heat exchanger, decompression means 9 and 12 are arranged in series, and the second indoor unit 3 is provided with a low-stage compressor 14 and a second four-way valve 36 capable of changing the discharge direction. As a result, the outdoor heat exchanger can be used for both the condenser and the evaporator. Therefore, the heat exchanger disposed in the room can be used for cooling as the evaporator and for heating as the condenser. In the case of a cycle using an ejector as shown in the configuration diagram of FIG. 15, the gas pipe 8 is extended to the second indoor heat exchanger 13 and connected to the anti-expansion valve side, and the first on-off valve is connected to the extension pipe. And a second on-off valve is provided in the suction flow pipe from the second indoor heat exchanger to the ejector 30, and the on-off valve is opened and closed in accordance with the switching of the four-way valve 35 on the heat source side. Heating is also possible. That is, when it is desired to air-condition the room, the already described operation is performed by closing the first on-off valve and opening the second on-off valve. On the other hand, when the heat source side is reversed, when the first on-off valve is opened and the second open / close valve is closed, the refrigerant is supplied to the heat exchangers 13 and 17 in parallel, and both heat exchangers serve as condensers. Operate. The ejector 30 is a pipe through which a refrigerant not related to any operation passes.

以上の様に、この発明に係る空気調和機は、容量調節可能な第１の圧縮機と、室外熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第２の室内送風機を具備する第１の室内ユニットと、第２の減圧手段、第２の室内熱交換器、風量調節可能な第２の室内送風機、容量調節可能な第２の圧縮機、を備えた第２の室内ユニットと、それらを接続する冷媒配管、で構成される空気調和機において、室内温湿度情報検出手段と、第２の室内熱交換器を流通する冷媒温度検出手段と、目標冷媒温度設定手段を備え、冷媒温度が目標冷媒温度となるように第２の室内送風機の送風量を制御するもので、この場合第１の圧縮機である高段圧縮機は室内温度の目標値である設定温度に近づけるように制御され、第２の圧縮機である低段圧縮機は室内湿度の目標値である設定湿度に近づけるように制御されるとともに第１の室内送風機である顕熱主体運転に使用される送風機は高段圧縮機に連動してリモコンなどで操作される運転モードに基づいて回転が設定され運転されるが、そのように設定がない場合の運転時には常に最大速度と設定すると良い。このような組み合わせの運転によりエネルギーが最も少ない状態で快適な室内空調を行うことが出来る。   As described above, the air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, an outdoor heat exchanger, an outdoor unit including an outdoor fan, a first decompression unit, and a first indoor heat. Exchanger, first indoor unit including second indoor blower, second decompression means, second indoor heat exchanger, second indoor blower with adjustable air volume, second compression with adjustable capacity In an air conditioner composed of a second indoor unit including an air conditioner and a refrigerant pipe connecting them, an indoor temperature / humidity information detecting means and a refrigerant temperature detecting means flowing through the second indoor heat exchanger And a target refrigerant temperature setting means for controlling the air flow rate of the second indoor fan so that the refrigerant temperature becomes the target refrigerant temperature. In this case, the high-stage compressor as the first compressor Is controlled so as to approach the set temperature which is the target value of the second The low-stage compressor, which is a compressor, is controlled so as to approach the set humidity, which is the target value of indoor humidity, and the blower used for sensible heat main operation, which is the first indoor fan, is linked to the high-stage compressor. The rotation is set based on the operation mode operated by a remote controller or the like, and the operation is performed. However, when there is no such setting, it is preferable to always set the maximum speed. Comfortable indoor air conditioning can be performed with the least amount of energy by such a combination of operations.

なお、このような冷凍サイクル装置を空調機に使用するが、湿度検出を行わないで湿度設定値を設けないような場合には高段圧縮機による温度制御や蒸発温度検出に基づく露点温度制御などにより行うことが出来る。   Such a refrigeration cycle apparatus is used for an air conditioner, but when humidity detection is not performed and a humidity set value is not provided, temperature control by a high-stage compressor, dew point temperature control based on evaporation temperature detection, etc. Can be done.

またこの発明は、第１の室内熱交換器表面温度が室内空気の露点温度より高くなるように運転されるので、広い表面積を有する熱交換器からのドレンがほとんど無くなりドレン処理構造が簡素化されるだけでなく、熱交換器の形態が露滴がフィン表面を流れ落ちるための傾きや溝構造等の対策や風速による飛び散り対策が簡素化され熱交換器配置や構造が自由に選択できるようになる。例えば炭酸ガス冷媒の場合横方向に配置するコルゲートフィンを使用しても露の流れの問題が無いため安心して使用できるなど更に効率の良い熱交換器構造が可能になる。また従来のエアコン室内機のようにくの字型の前側熱交換器と逆Ｖの字を形成する後側熱交換器における問題点は熱交換器に付着する露滴の処理と熱交換能力の増大であった。   In addition, since the first indoor heat exchanger surface temperature is operated to be higher than the dew point temperature of the room air, the drain from the heat exchanger having a large surface area is almost eliminated and the drain processing structure is simplified. In addition, the heat exchanger configuration simplifies countermeasures such as inclination and groove structure for dew droplets to flow down the fin surface, and measures against scattering due to wind speed, so that the heat exchanger arrangement and structure can be freely selected. . For example, in the case of a carbon dioxide refrigerant, even if corrugated fins arranged in the lateral direction are used, there is no problem of the flow of dew, so that a more efficient heat exchanger structure can be realized such that it can be used with confidence. In addition, the problem with the rear heat exchanger that forms an inverted V shape with the front heat exchanger in the shape of a dog like a conventional air conditioner indoor unit is that of the treatment of dew droplets adhering to the heat exchanger and the heat exchange capacity. It was an increase.

図２３は熱交換器性能を大幅に向上させたエアコンの室内機構造説明図である。従来技術の問題点に対し壁３７に固定した室内機ケーシング３８の吸気流側である前側熱交換器１０を顕熱主体の構成にすることにより上方が覆い被さるような一円形状の熱交換器にして背部側の途中まで露点温度より高い温度の冷媒を流す様に伝熱管の接続配置、即ち図１の第１の室内ユニット対応とする。これにより顕熱処理能力を大幅に増大させることが出来るとともに例え上部先端が垂れ下がる形状でも露つきの問題から開放されており問題にならない。一方後側熱交換器は前側より大幅に小型、即ち前側の１／３より少ない面積の潜熱処理のもので、例えば伝熱管や伝熱フィンは前側も後側も一体形状の如く同一形状とするが、後側の伝熱管に流れる冷媒の温度は露点温度より低いものが流れる冷媒サイクルの接続、即ち図１の第２の室内ユニット対応の熱交換器１３になる。これにより後側熱交換器だけドレンの処理しやすい形状としておけばよい。特にユニット３８を壁面に取り付ける部分である背面構造を利用すれば深溝形状のドレンパン４０が可能であるし、室外への廃水処理も簡単になる。なお前側のドレンパン３９は浅い溝形状のものを取り付けておくだけで室内高湿度時の一時的な貯水対策とすることが出来る。このようなほぼ一円形態の前側および後側の熱交換器が１つの横流ファンに下方を接近させ上方に大きな空間を設けた形状で取り囲む形態の室内ファンにすることで大幅な処理能力増大と効率向上が可能になる。また室内側のユニットを一つに纏め、且つ、送風機も一つにするなど構造の簡素化が行える。特に炭酸ガスを使用した冷凍サイクルにとってエジェクタを室内機本体の背面側、例えば吹出し口の背部などに設ける構造で更に効率向上がコンパクトの構造で纏め上げることが出来る。又従来冷媒配管７、８を据付け状態に応じて設けていたケーシング３８の背部の隙間にエジェクタ３０やこのエジェクタ使用時の冷房と暖房の回路切換を行う開閉弁４６やなどを収納することが出来、効率の良い、しかも自然冷媒を使用したエアコンが小形で簡単な構成で纏めることが出来る。   FIG. 23 is an explanatory view of the indoor unit structure of an air conditioner in which the heat exchanger performance is greatly improved. The conventional heat exchanger 10 which is the intake air flow side of the indoor unit casing 38 fixed to the wall 37 with respect to the problems of the prior art is configured as a sensible heat main body so that the upper part covers the upper part. Thus, the connection arrangement of the heat transfer tubes, that is, the first indoor unit shown in FIG. 1 is adapted so that the refrigerant having a temperature higher than the dew point temperature flows halfway along the back side. As a result, the sensible heat treatment ability can be greatly increased, and even the shape in which the upper tip hangs down is free from the problem of dew and does not become a problem. On the other hand, the rear heat exchanger is much smaller than the front side, that is, has a latent heat treatment less than 1/3 of the front side. For example, the heat transfer tubes and heat transfer fins have the same shape on both the front side and the rear side, such as an integrated shape. However, the refrigerant flowing in the rear heat transfer tube is connected to the refrigerant cycle in which the temperature lower than the dew point temperature, that is, the heat exchanger 13 corresponding to the second indoor unit in FIG. Thus, only the rear heat exchanger needs to have a shape that facilitates drainage. In particular, if a rear structure, which is a part for attaching the unit 38 to a wall surface, is used, a deep groove-shaped drain pan 40 is possible, and wastewater treatment to the outside becomes easy. The drain pan 39 on the front side can be used as a temporary water storage measure at high humidity indoors by simply attaching a shallow groove-shaped one. Such a substantially circular heat exchanger on the front side and rear side approaches a single cross-flow fan at the bottom and surrounds it with a shape in which a large space is provided above, thereby greatly increasing the processing capacity. Efficiency can be improved. In addition, the structure can be simplified by combining indoor units into one unit and using one blower. In particular, for a refrigeration cycle using carbon dioxide gas, a structure in which an ejector is provided on the back side of the indoor unit main body, for example, the back of the outlet, etc., can further improve the efficiency with a compact structure. In addition, the ejector 30 and the opening / closing valve 46 for switching between the cooling and heating circuits when the ejector is used can be accommodated in the gap at the back of the casing 38 where the refrigerant pipes 7 and 8 have been provided according to the installation state. An air conditioner that is efficient and uses natural refrigerant can be combined in a small and simple configuration.

またこの発明は、第１の圧縮機が目標温度と現在の温度との偏差に基づいて容量制御され、第２の圧縮機が目標湿度と現在の湿度との偏差に基づいて容量制御され、最も効率の良い個々の制御が可能になるだけでなく、総合的な室内熱負荷処理を冷凍サイクル全体で行うため効果的な装置が可能になる。 Further, according to the present invention, the capacity of the first compressor is controlled based on the deviation between the target temperature and the current temperature, and the capacity of the second compressor is controlled based on the deviation between the target humidity and the current humidity. Not only can individual control with high efficiency be possible, but an effective apparatus is possible because comprehensive indoor heat load processing is performed in the entire refrigeration cycle.

またこの発明は、第２の室内ユニットが、第３の減圧手段と、第３の室内熱交換器を備え、室内空気が前記第３の室内熱交換器を通過後に第２の室内熱交換器を通過するように構成されているので、冷房時に高段側、すなわち顕熱負荷の処理を有効に生かすことが出来、更に暖房負荷に対しても処理能力の向上と効率向上を可能にしている。
ある。 Further, according to the present invention, the second indoor unit includes a third decompression unit and a third indoor heat exchanger, and the second indoor heat exchanger after the indoor air passes through the third indoor heat exchanger. Since it is configured to pass through, it is possible to effectively utilize the processing of the high stage side, that is, the sensible heat load at the time of cooling, and further, it is possible to improve the processing capacity and the efficiency for the heating load. .
is there.

またこの発明は、冷媒回路に封入される冷媒が、二酸化炭素の様に異なる蒸発温度で利用できるエンタルピ差があるものにしたので、ハイドロフルオロカーボン系の冷媒を用いるよりも高効率で運転することができる。   In addition, since the refrigerant sealed in the refrigerant circuit has an enthalpy difference that can be used at different evaporation temperatures such as carbon dioxide, the refrigerant circuit can be operated with higher efficiency than using a hydrofluorocarbon refrigerant. it can.

又この発明の空気調和機は容量調節可能もしくは調節不可能な第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し冷媒を循環させる第１の冷媒サイクルと、第１の冷媒サイクルに第１の減圧手段および第１の負荷側熱交換器と並列に接続され第２の減圧手段、第２の負荷側熱交換器および第１の圧縮機とは独立に運転される第２の圧縮機に冷媒を循環させる第２の冷媒サイクルと、第１の負荷側熱交換器及び第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、第１の負荷側熱交換器の表面温度が空調領域の空気の露点温度より高くなり、第２の負荷側熱交換器の表面温度が空調領域の露点温度より低くなる様に、第１の負荷側熱交換器の熱処理能力を第２の負荷側熱交換器の熱処理能力より大きくしたものである。   In the air conditioner of the present invention, the first compressor, the heat source side heat exchanger, the first pressure reducing means, and the first load side heat exchanger, which are adjustable or not adjustable in capacity, are sequentially connected to circulate the refrigerant. The first refrigerant cycle, the first refrigerant cycle and the first decompression means and the first load side heat exchanger connected in parallel with the second decompression means, the second load side heat exchanger, and the first refrigerant cycle The second refrigerant cycle that circulates the refrigerant to the second compressor that is operated independently of the compressor, and the air that is air-conditioned in the first load-side heat exchanger and the second load-side heat exchange are the same. Load-side air blowing means that blows out to the air-conditioning region, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is The heat treatment capacity of the first load-side heat exchanger is set to the second load so that it is lower than the dew point temperature in the air conditioning area. It is made larger than the heat treatment capability of the heat exchanger.

又この発明の空気調和機は、容量調節可能な第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し冷媒を循環させる第１の冷媒サイクルと、第１の冷媒サイクルの前記第１の減圧手段であってこの第１の減圧手段と並列に第２の減圧手段および第２の負荷側熱交換器を接続し、第１の減圧手段の減圧時の駆動流による吸引を利用して前記第２の負荷側熱交換器からの冷媒圧力を昇圧するエジェクタと、第１の負荷側熱交換器及び前記第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、第１の負荷側熱交換器の表面温度が空調領域の空気の露点温度より高くなり、第２の負荷側熱交換器の表面温度が空調領域の露点温度より低くなる様に、第１の負荷側熱交換器の熱処理能力を第２の負荷側熱交換器の熱処理能力より大きくしたものである。   In the air conditioner of the present invention, the first compressor whose capacity is adjustable, the heat source side heat exchanger, the first pressure reducing means, and the first load side heat exchanger are sequentially connected to circulate the refrigerant. A first depressurization unit of the refrigerant cycle and the first refrigerant cycle, wherein a second depressurization unit and a second load-side heat exchanger are connected in parallel with the first depressurization unit, and the first depressurization unit An ejector for increasing the refrigerant pressure from the second load-side heat exchanger by using suction by the driving flow at the time of pressure reduction of the means, the first load-side heat exchanger, and the second load-side heat exchange. Load-side air blowing means for blowing air that has been air-conditioned to the same air-conditioning area, and the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning area, The heat of the first load-side heat exchanger is such that the surface temperature of the heat exchanger is lower than the dew point temperature of the air conditioning area. The physical capacity is made larger than the heat treatment capability of the second load-side heat exchanger.

又この発明の空気調和機は、容量調節可能な第１の圧縮機、室外熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、第２の減圧手段、第２の室内熱交換器、風量調節可能な第２の室内送風機、および第１の圧縮機とは独立に容量調節可能な第２の圧縮機、を具備する第２の室内ユニットと、室外ユニット、第１の室内ユニット、第２の室内ユニットを高圧側と低圧側をあわせて接続する冷媒配管と、を備え、第１の室内熱交換器表面温度が室内空気の露点温度より高くなるように運転されるものである。   The air conditioner of the present invention includes an outdoor unit having a capacity-adjustable first compressor, an outdoor heat exchanger, an outdoor blower, a first decompression means, a first indoor heat exchanger, a first The capacity can be adjusted independently of the first indoor unit including the indoor blower, the second decompression means, the second indoor heat exchanger, the second indoor blower capable of adjusting the air volume, and the first compressor. A second indoor unit comprising a second compressor, an outdoor unit, a first indoor unit, and a refrigerant pipe connecting the second indoor unit together on the high pressure side and the low pressure side, Are operated so that the surface temperature of the indoor heat exchanger becomes higher than the dew point temperature of the indoor air.

この発明は、容量調節可能な第１の圧縮機、室外熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、膨張動力回収手段、第２の減圧手段、第２の室内熱交換器、第２の室内送風機、第２の圧縮機を具備する第２の室内ユニットと、室外ユニット、第１の室内ユニット、第２の室内ユニットを高圧側液配管と低圧側ガス配管で接続し、高圧側の冷媒により膨張動力回収手段で回収された動力で第２の圧縮機を駆動するものである。   The present invention includes an outdoor unit including a first compressor with adjustable capacity, an outdoor heat exchanger, and an outdoor fan, a first decompression unit, a first indoor heat exchanger, and a first indoor fan. A first indoor unit, an expansion power recovery means, a second decompression means, a second indoor heat exchanger, a second indoor blower, a second indoor unit comprising a second compressor, an outdoor unit, The first indoor unit and the second indoor unit are connected by a high pressure side liquid pipe and a low pressure side gas pipe, and the second compressor is driven by the power recovered by the expansion power recovery means by the high pressure side refrigerant. is there.

この発明は、第１の減圧手段と第１の負荷側熱交換器、および第２の減圧手段と第２の負荷側熱交換器は、それぞれ１台以上設けられると共に、負荷側送風手段は第１の負荷側熱交換器および第２負荷側熱交換器に少なくとも１台設けられる。   In the present invention, one or more first decompression means and first load side heat exchanger, and two or more second decompression means and second load side heat exchangers are provided, respectively, and One load-side heat exchanger and at least one second load-side heat exchanger are provided.

この発明は、第１の負荷側熱交換器もしくは第１の室内熱交換器は表面温度が空調領域の空気の温度より１０度以上高い顕熱熱交換器であり、第２の負荷側熱交換器もしくは第２の室内熱交換器は、表面温度が空調領域の空気の温度より１０度以内である潜熱熱交換器である。   In the present invention, the first load-side heat exchanger or the first indoor heat exchanger is a sensible heat exchanger whose surface temperature is 10 degrees or more higher than the temperature of air in the air-conditioned region, and the second load-side heat exchange Or the second indoor heat exchanger is a latent heat exchanger whose surface temperature is within 10 degrees from the temperature of air in the air-conditioned area.

この発明は、第１の減圧手段および第１の負荷側熱交換器もしくは第１の室内側熱交換器と並列に、第３の減圧手段、第３の熱交換器を設け、室内空気が第３の熱交換器を通過後に第２の負荷側熱交換器もしくは第２の室内熱交換器を通過するように構成されている。   According to the present invention, a third pressure reducing means and a third heat exchanger are provided in parallel with the first pressure reducing means and the first load side heat exchanger or the first indoor side heat exchanger. After passing through the heat exchanger 3, the second load-side heat exchanger or the second indoor heat exchanger is passed.

この発明は、第２の負荷側熱交換器もしくは第２の室内側熱交換器の吸込み空気温度を下げる様に第２の負荷側熱交換器もしくは第２の室内側熱交換器の空気吸込み側に第１の負荷側熱交換器もしくは第１の室内側熱交換器とは異なる顕熱熱交換器を設けたものである。   The present invention provides an air suction side of the second load side heat exchanger or the second indoor side heat exchanger so as to lower the intake air temperature of the second load side heat exchanger or the second indoor side heat exchanger. Are provided with a sensible heat exchanger different from the first load side heat exchanger or the first indoor side heat exchanger.

この発明は、第２の負荷側熱交換器もしくは第２の室内側熱交換器の吹出し空気と第２の負荷側熱交換器もしくは第２の室内側熱交換器の吸込み空気との間で熱交換する全熱熱交換器、ヒートパイプのような顕熱熱交換器を第２の負荷側熱交換器もしくは第２の室内側熱交換器の空気吸込み側に設けたものである。   The present invention provides heat between the blowout air of the second load side heat exchanger or the second indoor side heat exchanger and the intake air of the second load side heat exchanger or the second indoor side heat exchanger. A total heat heat exchanger to be exchanged and a sensible heat exchanger such as a heat pipe are provided on the air suction side of the second load side heat exchanger or the second indoor side heat exchanger.

又この発明は第１の圧縮機の冷媒を吐出する側と吸込み側を切換可能に熱源側熱交換器および第１の負荷側熱交換器もしくは第１の室内側熱交換器との間に切換弁を設けたものである。又この発明は、第２の冷媒サイクルの第１の圧縮機側と熱源側熱交換器側とを切換可能に第２の切換弁を設けたものである。   Further, the present invention switches between the heat source side heat exchanger and the first load side heat exchanger or the first indoor side heat exchanger so that the refrigerant discharge side and the suction side of the first compressor can be switched. A valve is provided. Further, according to the present invention, a second switching valve is provided so as to be switchable between the first compressor side and the heat source side heat exchanger side of the second refrigerant cycle.

又この発明は、第１の冷媒サイクルの高圧側に液冷媒の高圧で駆動される膨張機を設け、第１の冷媒サイクルの低圧側に膨張機にて駆動される第２の圧縮機を設けたものである。   In the present invention, an expander driven by the high pressure of the liquid refrigerant is provided on the high pressure side of the first refrigerant cycle, and a second compressor driven by the expander is provided on the low pressure side of the first refrigerant cycle. It is a thing.

この発明は、第２の負荷側熱交換器もしくは第２の室内熱交換器を流通する冷媒の温度を検出する冷媒温度検出手段と、を備え、冷媒温度が目標冷媒温度となるように第２の負荷側熱交換器もしくは第２の室内熱交換器に送風する送風手段の送風量を制御するものである。   The present invention comprises refrigerant temperature detecting means for detecting the temperature of the refrigerant flowing through the second load-side heat exchanger or the second indoor heat exchanger, and the second temperature so that the refrigerant temperature becomes the target refrigerant temperature. The air flow rate of the air blowing means for blowing air to the load side heat exchanger or the second indoor heat exchanger is controlled.

又この発明は、第１の圧縮機が目標温度と空調領域の空気温度との偏差に基づいて容量制御され、第２の圧縮機が目標湿度と空調領域の空気湿度との偏差に基づいて容量制御される。また第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の吸込み空気が取入れ外気である。又第１の圧縮機の最大容量は、第２の圧縮機最大容量の処理熱量の３倍以上である。又第１の負荷側熱交換器もしくは第１の室内熱交換器の伝熱面積は、第２の負荷側熱交換器もしくは第２の室内熱交換器の伝熱面積の３倍以上である。   In the present invention, the capacity of the first compressor is controlled based on the deviation between the target temperature and the air temperature in the air-conditioned area, and the capacity of the second compressor is based on the deviation between the target humidity and the air humidity in the air-conditioned area. Be controlled. The intake air of the second load-side heat exchanger or the second indoor-side heat exchanger is intake outside air. Further, the maximum capacity of the first compressor is three times or more the processing heat amount of the second compressor maximum capacity. The heat transfer area of the first load-side heat exchanger or the first indoor heat exchanger is three times or more than the heat transfer area of the second load-side heat exchanger or the second indoor heat exchanger.

この発明は、第１の負荷側熱交換器もしくは第１の室内側熱交換器、および第２の負荷側熱交換器もしくは第２の室内側熱交換器を一つの箱体の内部に収納して同一の空調流域を空調するものである。   According to the present invention, the first load-side heat exchanger or the first indoor-side heat exchanger, and the second load-side heat exchanger or the second indoor-side heat exchanger are accommodated in one box. Air-conditioning the same air-conditioning basin.

又この発明は、減圧手段、第２の圧縮機、エジェクタ、開閉弁、等をユニットケーシングなどの箱体の内部に収納する。   Further, according to the present invention, the decompression means, the second compressor, the ejector, the on-off valve, and the like are accommodated in a box body such as a unit casing.

この発明の空気調和機は、低段と高段の異なる２つの蒸発圧力で同一の冷房負荷を処理する際に、低段側蒸発温度を消費電力が最小となるように風量調節するようにしたので、冷凍サイクル全体が高効率運転を行うことができる。   In the air conditioner of the present invention, when the same cooling load is processed with two different evaporating pressures of the low stage and the high stage, the air volume of the low stage side evaporating temperature is adjusted so that the power consumption is minimized. Thus, the entire refrigeration cycle can be operated with high efficiency.

また、この発明は、高段側熱交換器で空気を予冷した後に低段側熱交換器で冷却除湿するので、低段側の圧縮機容量比が高段側に対して小さくなり、トータルの消費電力を小さくすることができる。また、低段側での処理空気を取入れ外気とすることで、さらに消費電力を小さくすることができる。   In addition, since the present invention cools and dehumidifies in the low-stage heat exchanger after pre-cooling the air in the high-stage heat exchanger, the compressor capacity ratio on the low-stage side becomes smaller than that on the high-stage side, and the total Power consumption can be reduced. In addition, the power consumption can be further reduced by introducing the processing air on the lower stage side to be the outside air.

また、この発明は、冷媒を二酸化炭素としたので、ハイドロフルオロカーボン系の冷媒を用いるよりも高効率で運転することができる。   Further, since the present invention uses carbon dioxide as the refrigerant, it can be operated with higher efficiency than using a hydrofluorocarbon-based refrigerant.

この発明の実施の形態１を示す空気調和機の冷媒回路図である。It is a refrigerant circuit figure of the air conditioner which shows Embodiment 1 of this invention. この発明の実施の形態１の冷凍サイクル動作を示すＰ−ｈ線図である。It is a Ph diagram which shows the refrigerating cycle operation | movement of Embodiment 1 of this invention. この発明の実施の形態１における低段熱交換器の蒸発温度変化に対する顕熱比特性を示す特性図である。It is a characteristic view which shows the sensible heat ratio characteristic with respect to the evaporation temperature change of the low stage heat exchanger in Embodiment 1 of this invention. この発明の実施の形態１の低段蒸発温度に対する低段圧縮機入力変化を示す特性図である。It is a characteristic view which shows the low stage compressor input change with respect to the low stage evaporation temperature of Embodiment 1 of this invention. この発明の実施の形態１における空気調和機の制御動作を示すフローチャートである。It is a flowchart which shows the control operation of the air conditioner in Embodiment 1 of this invention. この発明の実施の形態１における低段蒸発温度とＣＯＰとの関係を示す特性図である。It is a characteristic view which shows the relationship between the low stage evaporation temperature and COP in Embodiment 1 of this invention. この発明の別の実施の形態における冷媒回路図である。It is a refrigerant circuit figure in another embodiment of this invention. この発明の実施の形態１における二酸化炭素冷媒を封入した際の冷凍サイクル動作を示すＰ−ｈ線図である。It is a Ph diagram which shows the refrigerating cycle operation | movement at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. この発明の実施の形態１における二酸化炭素冷媒を封入した際の冷凍サイクル動作を示すＰ−ｈ線図である。It is a Ph diagram which shows the refrigerating cycle operation | movement at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. この発明の実施の形態１における二酸化炭素冷媒を封入した際の別の冷媒回路図である。It is another refrigerant circuit figure at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. この発明の実施の形態１を示す外気処理状態を説明する空気線図の説明図である。It is explanatory drawing of the air diagram explaining the external air processing state which shows Embodiment 1 of this invention. この発明の実施の形態１の予冷・再熱構成を説明する説明図である。It is explanatory drawing explaining the pre-cooling and reheating structure of Embodiment 1 of this invention. この発明の実施の形態１における予冷・再熱構成の動作説明図である。It is operation | movement explanatory drawing of the pre-cooling and reheating structure in Embodiment 1 of this invention. この発明の実施の形態１のエジェクタの原理説明図である。It is principle explanatory drawing of the ejector of Embodiment 1 of this invention. この発明の実施の形態１におけるエジェクタを使用した冷凍サイクル構成図である。It is a refrigerating cycle block diagram using the ejector in Embodiment 1 of this invention. この発明の実施の形態１のエジェクタを使用した冷凍サイクルにおけるＰｈ線図である。It is a Ph diagram in the refrigerating cycle using the ejector of Embodiment 1 of this invention. この発明の実施の形態１におけるエジェクタを使用した冷凍サイクル動作説明図である。It is refrigerating cycle operation | movement explanatory drawing using the ejector in Embodiment 1 of this invention. この発明の実施の形態１におけるエジェクタを使用した別の冷凍サイクル構成図である。It is another refrigeration cycle block diagram using the ejector in Embodiment 1 of this invention. この発明の実施の形態１における空気調和機の冷凍サイクル構成の動作を示すＰ−ｈ線図である。It is a Ph diagram which shows operation | movement of the refrigerating cycle structure of the air conditioner in Embodiment 1 of this invention. この発明の実施の形態１におけるエジェクタを使用した部分的な冷凍サイクル構成説明図である。It is a partial refrigeration cycle structure explanatory view using the ejector in Embodiment 1 of this invention. この発明の実施の形態１における空気調和機の構造説明図である。It is structure explanatory drawing of the air conditioner in Embodiment 1 of this invention. この発明の実施の形態１における別の冷凍サイクル構成図である。It is another refrigeration cycle block diagram in Embodiment 1 of this invention. この発明の実施の形態１における空気調和機の別の構成図である。It is another block diagram of the air conditioner in Embodiment 1 of this invention.

符号の説明Explanation of symbols

１ 室外ユニット、２ 第１の室内ユニット、３ 第２の室内ユニット、４ 高段圧縮機、５ 室外熱交換器、６ 室外送風機、７ 冷媒配管（液管）、８ 冷媒配管（ガス管）、９ 第１減圧手段、１０ 第１室内熱交換器、１１ 第１室内送風機、１２ 第２減圧手段、１３ 第２室内熱交換器、１４ 低段圧縮機、１５ 第２室内送風機、１６ 第３減圧手段、１７ 第３室内熱交換器、２１ 膨張動力回収手段、２２ 膨張部、２３ 圧縮部、３０ エジェクタ、３１ ノズル、３２ 混合部、３３ ディフューザ部、３４ 室内機、３５ 第１の四方弁、３６ 第２の四方弁、３７ 壁、３８ ケーシング、３９ 顕熱側ドレンパン、４０ 潜熱側ドレンパン、４１ 横流ファン、４２ 吹出口、４３ 風案内板、４４ 顕熱側伝熱管、４５ 潜熱側伝熱管、４６ エジェクタ、５０ 天井、５１ 天井埋め込みカセット、５２ 送風機、５３ フィルタ、５４ ドレンパン、５５ 開閉弁などの収納箱、５６ 全熱熱交換器。   1 outdoor unit, 2 first indoor unit, 2nd indoor unit, 4 high stage compressor, 5 outdoor heat exchanger, 6 outdoor blower, 7 refrigerant pipe (liquid pipe), 8 refrigerant pipe (gas pipe), DESCRIPTION OF SYMBOLS 9 1st pressure reduction means, 10 1st indoor heat exchanger, 11 1st indoor fan, 12 2nd pressure reduction means, 13 2nd indoor heat exchanger, 14 Low stage compressor, 15 2nd indoor fan, 16 3rd pressure reduction Means, 17 third indoor heat exchanger, 21 expansion power recovery means, 22 expansion section, 23 compression section, 30 ejector, 31 nozzle, 32 mixing section, 33 diffuser section, 34 indoor unit, 35 first four-way valve, 36 Second four-way valve, 37 wall, 38 casing, 39 sensible heat side drain pan, 40 latent heat side drain pan, 41 cross flow fan, 42 outlet, 43 wind guide plate, 44 sensible heat side heat transfer tube, 45 latent heat side heat transfer tube 46 ejector 50 ceiling, 51 ceiling cassette, 52 a blower, 53 a filter, 54 a drain pan, storage box, such as 55-off valve, 56 total heat exchanger.

Claims (24)

冷媒を吐出する第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し前記冷媒を循環させる第１の冷媒サイクルと、前記第１の冷媒サイクルに前記第１の減圧手段および前記第１の負荷側熱交換器と並列に接続される第２の減圧手段、第２の負荷側熱交換器および前記第１の圧縮機とは独立に運転される第２の圧縮機を有する第２の冷媒サイクルと、前記第１の負荷側熱交換器及び前記第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、前記第１の負荷側熱交換器の温度が前記空調領域の空気の露点温度より高くなり、前記第２の負荷側熱交換器の温度が前記空調領域の露点温度より低くなる様に、前記第１の負荷側熱交換器の熱処理能力を前記第２の負荷側熱交換器の熱処理能力より大きくしたことを特徴とする空気調和機。 A first refrigerant cycle for discharging the refrigerant, a heat source side heat exchanger, a first pressure reducing means, a first load side heat exchanger in order and circulating the refrigerant; Independently of the second pressure reducing means, the second load side heat exchanger and the first compressor connected in parallel with the first pressure reducing means and the first load side heat exchanger in the refrigerant cycle A load that blows out air conditioned by the second refrigerant cycle having the second compressor to be operated, the first load-side heat exchanger, and the second load-side heat exchange to the same air-conditioning region Side air blowing means, wherein the temperature of the first load-side heat exchanger is higher than the dew point temperature of the air in the air-conditioned area, and the temperature of the second load-side heat exchanger is the dew point temperature of the air-conditioned area. The heat treatment capacity of the first load-side heat exchanger is reduced to the second load-side heat exchange so as to be lower. An air conditioner characterized by being larger than the vessels of heat treatment capacity. 容量調節可能な第１の圧縮機、熱源側熱交換器、第１の減圧手段、第１の負荷側熱交換器を順次接続し冷媒を循環させる第１の冷媒サイクルと、前記第１の冷媒サイクルの前記第１の減圧手段であってこの第１の減圧手段と並列に第２の減圧手段および第２の負荷側熱交換器を接続し、前記第１の減圧手段の減圧時の駆動流による吸引を利用して前記第２の負荷側熱交換器からの冷媒圧力を昇圧する昇圧手段と、前記第１の負荷側熱交換器及び前記第２の負荷側熱交換にて空調された空気を同一の空調領域に吹出す負荷側送風手段と、を備え、前記第１の負荷側熱交換器の表面温度が前記空調領域の空気の露点温度より高くなり、前記第２の負荷側熱交換器の表面温度が前記空調領域の露点温度より低くなる様に、前記第１の負荷側熱交換器の熱処理能力を前記第２の負荷側熱交換器の熱処理能力より大きくしたことを特徴とする空気調和機。 A first refrigerant cycle in which the capacity is adjustable, a first heat source side heat exchanger, a first pressure reducing means, a first load side heat exchanger and a first refrigerant cycle for circulating the refrigerant; and the first refrigerant A second decompression unit and a second load side heat exchanger connected to the first decompression unit in parallel with the first decompression unit, and a driving flow during decompression of the first decompression unit Boosting means for boosting the refrigerant pressure from the second load-side heat exchanger using suction by the air, air conditioned by the first load-side heat exchanger and the second load-side heat exchange Load-side air blowing means for blowing out air to the same air-conditioning region, and the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the second load-side heat exchange The first load-side heat exchanger so that the surface temperature of the cooler is lower than the dew point temperature of the air conditioning region Air conditioner, wherein a heat treatment capacity and greater than the second heat treatment capacity of the load-side heat exchanger. 前記第１の減圧手段と前記第１の負荷側熱交換器、および前記第２の減圧手段と前記第２の負荷側熱交換器は、それぞれ１台以上設けられると共に、前記負荷側送風手段は前記第１の負荷側熱交換器および前記第２負荷側熱交換器に少なくとも１台設けられることを特徴とする請求項１又は２記載の空気調和機。 One or more each of the first pressure reducing means and the first load side heat exchanger, and the second pressure reducing means and the second load side heat exchanger are provided, and the load side blowing means is The air conditioner according to claim 1 or 2, wherein at least one of the first load side heat exchanger and the second load side heat exchanger is provided. 容量調節可能な第１の圧縮機、熱源側熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、第２の減圧手段、第２の室内熱交換器、風量調節可能な第２の室内送風機、および前記第１の圧縮機とは独立に容量調節可能な第２の圧縮機、を具備する第２の室内ユニットと、前記室外ユニット、前記第１の室内ユニット、前記第２の室内ユニットを接続する冷媒配管と、を備え、前記第１の室内熱交換器表面温度が室内空気の露点温度より高くなるように運転されることを特徴とする空気調和機。 A first compressor having a capacity-adjustable first compressor, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, and a first indoor fan. An indoor unit, a second decompression means, a second indoor heat exchanger, a second indoor blower capable of adjusting the air volume, and a second compressor capable of adjusting the capacity independently of the first compressor, A second indoor unit, and a refrigerant pipe connecting the outdoor unit, the first indoor unit, and the second indoor unit, wherein the first indoor heat exchanger surface temperature is the room air temperature. An air conditioner that is operated so as to be higher than a dew point temperature. 容量調節可能な第１の圧縮機、熱源側熱交換器、室外送風機を具備する室外ユニットと、第１の減圧手段、第１の室内熱交換器、第１の室内送風機を具備する第１の室内ユニットと、膨張動力回収手段、第２の減圧手段、第２の室内熱交換器、第２の室内送風機、第２の圧縮機を具備する第２の室内ユニットと、前記室外ユニット、前記第１の室内ユニット、前記第２の室内ユニットを高圧側液配管と低圧側ガス配管で接続し、高圧側の冷媒により前記膨張動力回収手段で回収された動力で第２の圧縮機を駆動することを特徴とする空気調和機。 A first compressor having a capacity-adjustable first compressor, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, and a first indoor fan. An indoor unit, an expansion power recovery means, a second decompression means, a second indoor heat exchanger, a second indoor blower, a second indoor unit comprising a second compressor, the outdoor unit, the first 1 indoor unit and the second indoor unit are connected by a high pressure side liquid pipe and a low pressure side gas pipe, and the second compressor is driven by the power recovered by the expansion power recovery means by the high pressure side refrigerant. Air conditioner characterized by. 前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器の表面温度を計測し、この表面温度が効率の良い目標温度になるように負荷側送風手段もしくは第２の室内送風機、第２の圧縮機および第２の減圧手段の少なくとも１つを制御することを特徴とする請求項１乃至５のいずれかに記載の記載の空気調和機。 The surface temperature of the second load-side heat exchanger or the second indoor heat exchanger is measured, and the load-side blowing means or the second indoor fan, The air conditioner according to any one of claims 1 to 5, wherein at least one of the second compressor and the second pressure reducing means is controlled. 前記第１の圧縮機が目標温度と空調領域の空気温度との偏差に基づいて制御され、前記第２の圧縮機もしくは前記昇圧手段が目標湿度と空調領域の空気湿度との偏差に基づいて制御されるもしくは前記第２の負荷側熱交換器又は前記第２の室内熱交換器の表面温度と前記空調領域の空気温度との偏差に基づいて制御される、もしくは前記第２の負荷側熱交換器又は前記第２の室内熱交換器の表面温度と設定された目標値との差に基づいて制御されることを特徴とする請求項１乃至５のいずれかに記載の記載の空気調和機。 The first compressor is controlled based on a deviation between the target temperature and the air temperature in the air-conditioning area, and the second compressor or the booster is controlled based on a deviation between the target humidity and the air humidity in the air-conditioning area. Or controlled based on a deviation between the surface temperature of the second load-side heat exchanger or the second indoor heat exchanger and the air temperature of the air-conditioned area, or the second load-side heat exchange The air conditioner according to any one of claims 1 to 5, wherein the air conditioner is controlled based on a difference between a surface temperature of the heater or the second indoor heat exchanger and a set target value. 前記第１の負荷側熱交換器もしくは前記第１の室内熱交換器の伝熱面積を前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器の伝熱面積より大きくしたことを特徴とする請求項１乃至７のいずれかに記載の空気調和機。 The heat transfer area of the first load side heat exchanger or the first indoor heat exchanger is made larger than the heat transfer area of the second load side heat exchanger or the second indoor heat exchanger. The air conditioner according to any one of claims 1 to 7, wherein 前記第１の負荷側熱交換器もしくは前記第１の室内熱交換器は表面温度と空気の温度との差が１０度程度以内となる顕熱熱交換器であり、前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器は、表面温度が空気温度より１０度程度以上低い潜熱熱交換器であることを特徴とする請求項１乃至８のいずれかに記載の空気調和機。 The first load-side heat exchanger or the first indoor heat exchanger is a sensible heat exchanger in which the difference between the surface temperature and the air temperature is within about 10 degrees, and the second load-side heat The air conditioner according to any one of claims 1 to 8, wherein the exchanger or the second indoor heat exchanger is a latent heat exchanger whose surface temperature is lower by about 10 degrees or more than the air temperature. 前記第１の減圧手段および前記第１の負荷側熱交換器もしくは前記第１の室内側熱交換器と並列に、第３の減圧手段と第３の熱交換器を設け、室内空気が前記第３の熱交換器を通過後に前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器を通過するように構成されていることを特徴とする請求項１乃至９記載の空気調和機。 A third pressure reducing means and a third heat exchanger are provided in parallel with the first pressure reducing means and the first load side heat exchanger or the first indoor side heat exchanger, and indoor air is supplied to the first pressure reducing means and the first indoor heat exchanger. 10. The air conditioner according to claim 1, wherein the air conditioner is configured to pass through the second load-side heat exchanger or the second indoor heat exchanger after passing through the heat exchanger 3. . 前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の吸込み空気温度を下げる様に前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の空気吸込み側に前記第１の負荷側熱交換器を設けたことを特徴とする請求項１乃至９のいずれかに記載の空気調和機。 The air intake side of the second load side heat exchanger or the second indoor side heat exchanger so as to lower the intake air temperature of the second load side heat exchanger or the second indoor side heat exchanger. The air conditioner according to any one of claims 1 to 9, wherein the first load side heat exchanger is provided in the air conditioner. 前記第２の負荷側熱交換器もしくは第２の室内側熱交換器の吹出し空気と前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の吸込み空気との間で熱交換する全熱熱交換器、ヒートパイプのような別の独立した顕熱熱交換器を前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の空気吸込み側に設けたことを特徴とする請求項１乃至９のいずれかに記載の空気調和機。 Heat exchange between the blown air of the second load-side heat exchanger or the second indoor heat exchanger and the intake air of the second load-side heat exchanger or the second indoor heat exchanger And a separate independent sensible heat exchanger such as a heat pipe is provided on the air intake side of the second load side heat exchanger or the second indoor side heat exchanger. The air conditioner according to any one of claims 1 to 9, wherein 前記第１の圧縮機の冷媒を吐出する側と吸込み側を切換可能に前記熱源側熱交換器および前記第１の負荷側熱交換器もしくは第１の室内側熱交換器との間に切換弁を設けたことを特徴とする請求項１乃至１２のいずれかに記載の空気調和機。 A switching valve between the heat source side heat exchanger and the first load side heat exchanger or the first indoor side heat exchanger so that the refrigerant discharge side and the suction side of the first compressor can be switched. The air conditioner according to claim 1, wherein the air conditioner is provided. 前記第２の冷媒サイクルもしくは第２の室内ユニットの前記第１の圧縮機側冷媒配管と前記熱源側熱交換器側冷媒配管とを切換可能に第２の切換弁もしくは開平弁を設けたバイパス配管を設けたことを特徴とする請求項１乃至１３のいずれかに記載の空気調和機。 Bypass piping provided with a second switching valve or square root valve so as to switch between the first compressor side refrigerant piping and the heat source side heat exchanger side refrigerant piping of the second refrigerant cycle or the second indoor unit. The air conditioner according to claim 1, wherein the air conditioner is provided. 前記第１の冷媒サイクルの高圧側もしくは室外ユニットからの高圧側の冷媒配管に液冷媒の高圧で駆動される膨張機を接続し、前記第１の冷媒サイクルの低圧側に前記膨張機にて駆動される第２の圧縮機を接続したことを特徴とする請求項１又は４又は５又は８に記載の空気調和機。 An expander driven by a high pressure of liquid refrigerant is connected to a refrigerant pipe on the high pressure side of the first refrigerant cycle or a high pressure side from the outdoor unit, and driven by the expander on the low pressure side of the first refrigerant cycle. The air conditioner according to claim 1, 4, 5, or 8, wherein the second compressor is connected. 前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器を流通する冷媒の温度を検出する冷媒温度検出手段と、を備え、前記冷媒温度が目標冷媒温度となるように前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器に送風する送風手段の送風量を制御することを特徴とする請求項１乃至１５のいずれかに記載の空気調和機。 Refrigerant temperature detecting means for detecting the temperature of the refrigerant flowing through the second load-side heat exchanger or the second indoor heat exchanger, and the second temperature so that the refrigerant temperature becomes a target refrigerant temperature. The air conditioner according to any one of claims 1 to 15, wherein a blower amount of a blower means for blowing air to the load side heat exchanger or the second indoor heat exchanger is controlled. 前記第２の減圧手段が目標湿度と空調領域の空気湿度との偏差もしくは前記第２の負荷側熱交換器の表面温度と前記空調領域の空気温度との偏差に基づいて開度制御されることを特徴とする請求項１乃至請求項１５のいずれかに記載の記載の空気調和機。 The degree of opening of the second pressure reducing means is controlled based on a deviation between a target humidity and an air humidity in the air conditioning area or a deviation between a surface temperature of the second load-side heat exchanger and the air temperature in the air conditioning area. The air conditioner according to any one of claims 1 to 15, wherein: 前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器の吸込み空気が外気であることを特徴とする請求項１乃至１５のいずれかに記載の空気調和機。 The air conditioner according to any one of claims 1 to 15, wherein the intake air of the second load-side heat exchanger or the second indoor-side heat exchanger is outside air. 前記第１の圧縮機の最大容量は、前記第２の圧縮機最大容量の処理熱量の３倍以上であることを特徴とする請求項１乃至９のいずれかに記載の空気調和機。 10. The air conditioner according to claim 1, wherein a maximum capacity of the first compressor is three times or more of a heat of treatment of the second compressor maximum capacity. 第１の負荷側熱交換器もしくは前記第１の室内熱交換器の伝熱面積は、前記第２の負荷側熱交換器もしくは前記第２の室内熱交換器の伝熱面積の３倍以上であることを特徴とする請求項１乃至９のいずれかに記載の空気調和機。 The heat transfer area of the first load-side heat exchanger or the first indoor heat exchanger is at least three times the heat transfer area of the second load-side heat exchanger or the second indoor heat exchanger. The air conditioner according to claim 1, wherein the air conditioner is provided. 前記第１の減圧手段、および前記第２の減圧手段は、その一端が接続されている熱交換器出口の冷媒過熱度が所定値となるように開度調節されることを特徴とする請求項１乃至２０のいずれかに記載の空気調和機。 2. The opening degree of the first decompression unit and the second decompression unit is adjusted so that a refrigerant superheat degree at a heat exchanger outlet connected to one end of the first decompression unit and the second decompression unit becomes a predetermined value. The air conditioner according to any one of 1 to 20. 前記冷媒回路に封入される冷媒が、二酸化炭素のような異なる蒸発圧力で蒸発機て利用できるエンタルピ差に差があるものであることを特徴とする請求項１乃至２０のいずれかに記載の空気調和機。 The air according to any one of claims 1 to 20, wherein the refrigerant sealed in the refrigerant circuit has a difference in enthalpy difference that can be used by an evaporator at different evaporation pressures such as carbon dioxide. Harmony machine. 前記第１の負荷側熱交換器もしくは前記第１の室内側熱交換器、および前記第２の負荷側熱交換器もしくは前記第２の室内側熱交換器を一つの箱体の内部に収納して同一の空調領域を空調することを特徴とする請求項１乃至２２のいずれかに記載の空気調和機。 The first load-side heat exchanger or the first indoor-side heat exchanger, and the second load-side heat exchanger or the second indoor-side heat exchanger are accommodated in one box. The air conditioner according to any one of claims 1 to 22, wherein the same air-conditioning region is air-conditioned. 前記第２の減圧手段、前記第２の圧縮機、前記昇圧手段、および冷媒の流を開閉する冷媒配管に設けられた開閉弁の少なくとも一つを前記箱体の内部に収納することを特徴とする請求項２３記載の空気調和機。 At least one of the second pressure reducing means, the second compressor, the pressure increasing means, and an on-off valve provided in a refrigerant pipe for opening and closing a refrigerant flow is housed in the box. The air conditioner according to claim 23.
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Cited By (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2009109064A (en) * 2007-10-29 2009-05-21 Hitachi Appliances Inc Air conditioner
JP2009180492A (en) * 2008-02-01 2009-08-13 Daikin Ind Ltd Dehumidifying unit and air conditioner
JP2009234387A (en) * 2008-03-26 2009-10-15 Calsonic Kansei Corp Air conditioner
JP2010038432A (en) * 2008-08-04 2010-02-18 Calsonic Kansei Corp Air conditioning system and control method of air-conditioning system
JP2011094840A (en) * 2009-10-28 2011-05-12 Corona Corp Heat pump device
JP2013204899A (en) * 2012-03-28 2013-10-07 Mitsubishi Electric Corp Air conditioning system
KR101505930B1 (en) 2014-12-30 2015-03-25 주식회사 나우이엘 Apparatus and method for controlling dehumidifier of ceiling reclamation type
CN107726554A (en) * 2017-09-19 2018-02-23 青岛海尔空调电子有限公司 A kind of multi-connected machine comfort level balance control method and system
CN110749064A (en) * 2019-10-31 2020-02-04 广东美的制冷设备有限公司 Air conditioner control method and system and air conditioner

Citations (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS4836460U (en) * 1971-09-03 1973-05-01
JPS5676957U (en) * 1979-11-20 1981-06-23
JPS5724428U (en) * 1980-07-17 1982-02-08
JP2002022298A (en) * 2000-07-04 2002-01-23 Matsushita Electric Ind Co Ltd Refrigeration cycle device and method for controlling the same
JP2002188841A (en) * 2000-12-18 2002-07-05 Mitsubishi Electric Corp Outdoor air treating unit
JP2002243306A (en) * 2001-02-13 2002-08-28 Mitsubishi Electric Corp Air conditioner
JP2003042582A (en) * 2002-07-18 2003-02-13 Mitsubishi Electric Corp Freezing cycle device
JP2003294296A (en) * 2002-03-29 2003-10-15 Daikin Ind Ltd Air conditioner
JP2004150750A (en) * 2002-10-31 2004-05-27 Matsushita Electric Ind Co Ltd Method for determining pressure of high pressure cooling medium in refrigerating cycle device
JP2005009725A (en) * 2003-06-18 2005-01-13 Sanyo Electric Co Ltd Air conditioner and control method of air conditioner
JP2006317012A (en) * 2005-03-07 2006-11-24 Kiyoshi Yanagimachi Air conditioner

Patent Citations (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS4836460U (en) * 1971-09-03 1973-05-01
JPS5676957U (en) * 1979-11-20 1981-06-23
JPS5724428U (en) * 1980-07-17 1982-02-08
JP2002022298A (en) * 2000-07-04 2002-01-23 Matsushita Electric Ind Co Ltd Refrigeration cycle device and method for controlling the same
JP2002188841A (en) * 2000-12-18 2002-07-05 Mitsubishi Electric Corp Outdoor air treating unit
JP2002243306A (en) * 2001-02-13 2002-08-28 Mitsubishi Electric Corp Air conditioner
JP2003294296A (en) * 2002-03-29 2003-10-15 Daikin Ind Ltd Air conditioner
JP2003042582A (en) * 2002-07-18 2003-02-13 Mitsubishi Electric Corp Freezing cycle device
JP2004150750A (en) * 2002-10-31 2004-05-27 Matsushita Electric Ind Co Ltd Method for determining pressure of high pressure cooling medium in refrigerating cycle device
JP2005009725A (en) * 2003-06-18 2005-01-13 Sanyo Electric Co Ltd Air conditioner and control method of air conditioner
JP2006317012A (en) * 2005-03-07 2006-11-24 Kiyoshi Yanagimachi Air conditioner

Cited By (10)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2009109064A (en) * 2007-10-29 2009-05-21 Hitachi Appliances Inc Air conditioner
JP2009180492A (en) * 2008-02-01 2009-08-13 Daikin Ind Ltd Dehumidifying unit and air conditioner
JP2009234387A (en) * 2008-03-26 2009-10-15 Calsonic Kansei Corp Air conditioner
JP2010038432A (en) * 2008-08-04 2010-02-18 Calsonic Kansei Corp Air conditioning system and control method of air-conditioning system
JP2011094840A (en) * 2009-10-28 2011-05-12 Corona Corp Heat pump device
JP2013204899A (en) * 2012-03-28 2013-10-07 Mitsubishi Electric Corp Air conditioning system
KR101505930B1 (en) 2014-12-30 2015-03-25 주식회사 나우이엘 Apparatus and method for controlling dehumidifier of ceiling reclamation type
CN107726554A (en) * 2017-09-19 2018-02-23 青岛海尔空调电子有限公司 A kind of multi-connected machine comfort level balance control method and system
CN107726554B (en) * 2017-09-19 2020-01-17 青岛海尔空调电子有限公司 Multi-split comfort level balance control method and system
CN110749064A (en) * 2019-10-31 2020-02-04 广东美的制冷设备有限公司 Air conditioner control method and system and air conditioner

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