JP4651627B2 - Refrigeration air conditioner - Google Patents

Refrigeration air conditioner Download PDF

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JP4651627B2
JP4651627B2 JP2007009995A JP2007009995A JP4651627B2 JP 4651627 B2 JP4651627 B2 JP 4651627B2 JP 2007009995 A JP2007009995 A JP 2007009995A JP 2007009995 A JP2007009995 A JP 2007009995A JP 4651627 B2 JP4651627 B2 JP 4651627B2
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heat exchanger
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compressor
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JP2008175476A (en
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多佳志 岡崎
史武 畝崎
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三菱電機株式会社
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この発明は、冷凍空調装置に関するものであり、特に水・ブラインなどの液媒体を加熱・冷却することにより冷温熱を負荷側に供給する冷凍空調装置に関するものである。   The present invention relates to a refrigeration air conditioner, and more particularly to a refrigeration air conditioner that supplies cold and hot heat to a load side by heating and cooling a liquid medium such as water and brine.
冷温熱を供給する冷凍空調装置の例として、例えば「室外に配置され冷却を外気により行なう凝縮器と、該凝縮器より高低差のある低位置に設けた蒸発器を含み、サーモサイフォン冷凍サイクル運転又は冷媒圧縮強制循環冷凍サイクル運転に切り替え可能に構成された冷凍サイクルを備えた冷却器において、単一の凝縮器に対して前記冷凍サイクルを複数段並列に配置し、複数段の冷凍サイクルの蒸発器を蒸発温度を1段目から順に低く設定してなる蒸発器で構成し、冷却負荷配管を前記冷凍サイクルの蒸発器に1段目から順に直列に接続してなる…」というものが提案されている(例えば特許文献1参照)。   As an example of a refrigerating and air-conditioning apparatus that supplies cold and hot heat, for example, a thermosiphon refrigeration cycle operation including a condenser that is arranged outside and performs cooling by outside air and an evaporator that is provided at a low position with a difference in elevation from the condenser Alternatively, in a cooler having a refrigeration cycle configured to be able to be switched to refrigerant compression forced circulation refrigeration cycle operation, the refrigeration cycle is arranged in a plurality of stages in parallel with respect to a single condenser, and evaporation of a plurality of refrigeration cycles The evaporator is composed of an evaporator in which the evaporating temperature is set in order lower from the first stage, and the cooling load pipe is connected in series from the first stage to the evaporator of the refrigeration cycle. (For example, refer to Patent Document 1).
特開2006−329601号公報(要約、1−10頁、図1)JP 2006-329601 A (Abstract, page 1-10, FIG. 1)
上記の特許文献1において提案されている冷凍空調装置は、温度差の大きい冷水の冷却に対して、複数の冷凍サイクルの蒸発器で段階的に冷却する場合に、各冷凍サイクルの蒸発器の蒸発温度を1段目から順に低く設定することで、高効率運転を行うとしているが、以下のような問題点があった。この冷凍空調装置では、各冷凍サイクルの圧縮機の容量をインバータにより制御するとしているが、外気温度に対応して圧縮機容量を低下するという記載や、冷媒圧力に基づいて圧縮機容量を制御するという記載があるのみで、様々な運転条件に対して、各圧縮機の容量をどのように制御すれば高効率な運転を実現できるかが示されておらず、結果として、運転条件に対応した高効率運転を実施できず、運転効率が低下するという問題点があった。   The refrigeration air conditioner proposed in the above-mentioned Patent Document 1 evaporates the evaporator of each refrigeration cycle when cooling the chilled water having a large temperature difference in stages with the evaporators of a plurality of refrigeration cycles. Although the high-efficiency operation is performed by setting the temperature sequentially lower from the first stage, there are the following problems. In this refrigeration air conditioner, the capacity of the compressor of each refrigeration cycle is controlled by an inverter. However, the compressor capacity is controlled based on the description that the compressor capacity is reduced according to the outside air temperature or the refrigerant pressure. However, there is no indication of how to control the capacity of each compressor for various operating conditions to achieve high-efficiency operation. There was a problem that high-efficiency operation could not be performed and the operation efficiency was lowered.
この発明は、以上の課題に鑑み、複数の冷凍サイクルで構成される冷凍空調装置において、運転条件に対応して圧縮機の運転容量制御を適切に行うことにより高効率の冷凍空調装置を得ることを目的とする。   In view of the above problems, the present invention obtains a highly efficient refrigeration air conditioner by appropriately controlling the operating capacity of the compressor in accordance with the operating conditions in the refrigeration air conditioner configured by a plurality of refrigeration cycles. With the goal.
本発明に係る冷凍空調装置は、運転容量が可変である圧縮機と、熱源側熱交換器と、減圧装置と、負荷側熱交換器とを環状に接続して構成される冷凍サイクルを複数備え、各冷凍サイクルの負荷側熱交換器において負荷側熱媒体を冷却又は加熱し、冷温熱を供給するとともに、負荷側熱媒体の流路が各冷凍サイクルの負荷側熱交換器を直列に流れるように構成され、各冷凍サイクルの負荷側熱交換器における負荷側熱媒体の流入温度及び流出温度をそれぞれ計測する温度センサと、負荷側熱交換器が負荷側熱媒体の流路の下流側に接続される冷凍サイクルの圧縮機容量を、各冷凍サイクルの圧縮機運転容量が負荷側熱交換器の容量又は圧縮機の定格容量に比例するようにして決定される圧縮機運転容量よりも大きく、かつ、負荷側熱交換器が負荷側熱媒体の流路の下流側に接続される冷凍サイクルの負荷側熱媒体の流入、流出の温度差が、各冷凍サイクルに搭載される負荷側熱交換器の容量又は圧縮機の定格容量に比例するようにして決定される温度差よりも小さくなるように、圧縮機運転容量を制御する制御装置とを備えたものである。 A refrigerating and air-conditioning apparatus according to the present invention includes a plurality of refrigeration cycles configured by annularly connecting a compressor having a variable operating capacity, a heat source side heat exchanger, a pressure reducing device, and a load side heat exchanger. In the load-side heat exchanger of each refrigeration cycle, the load-side heat medium is cooled or heated to supply cold / hot heat, and the flow path of the load-side heat medium flows in series through the load-side heat exchanger of each refrigeration cycle. A temperature sensor that measures the inflow temperature and the outflow temperature of the load-side heat medium in the load-side heat exchanger of each refrigeration cycle, and the load-side heat exchanger is connected to the downstream side of the flow path of the load-side heat medium The compressor capacity of each refrigeration cycle is greater than the compressor operating capacity determined such that the compressor operating capacity of each refrigeration cycle is proportional to the capacity of the load side heat exchanger or the rated capacity of the compressor, and , the load-side heat exchanger Inflow of load-side heat medium of the refrigerating cycle that is connected to the downstream side of the flow path of the load-side heat medium, the temperature difference between the outflow, the rated capacity of the capacity or compressor load-side heat exchanger mounted in the refrigeration cycle And a control device for controlling the compressor operating capacity so as to be smaller than the temperature difference determined in proportion to.
この発明に係る冷凍空調装置は、各冷凍サイクルがバランスよく熱負荷を賄えるように圧縮機容量を制御することで装置全体の運転効率を高め、高効率の冷凍空調装置を得ることができる。   The refrigerating and air-conditioning apparatus according to the present invention can increase the operation efficiency of the entire apparatus by controlling the compressor capacity so that each refrigeration cycle can cover the heat load in a well-balanced manner, thereby obtaining a highly efficient refrigerating and air-conditioning apparatus.
実施の形態1.
以下この発明の実施の形態1を図1に示す。図1はこの発明の冷凍空調装置の回路図である。冷凍空調装置である熱源機1内には、同一回路構成の冷凍サイクル2a、2bが搭載されている。冷凍サイクル2aには圧縮機3a、四方弁4a、熱源側熱交換器である空気熱交換器5a、逆止弁6a〜6d、過冷却熱交換器7a、減圧装置である主膨張弁8a、負荷側熱交換器である水熱交換器9a、第二の減圧装置であるバイパス膨張弁10aが内蔵されており、図示されるように環状に接続され冷媒回路を構成する。また、冷凍サイクル2bにおいても、同様にして、圧縮機3b、四方弁4b、熱源側熱交換器である空気熱交換器5b、逆止弁6e〜6h、過冷却熱交換器7b、減圧装置である主膨張弁8b、負荷側熱交換器である水熱交換器9b、第二の減圧装置であるバイパス膨張弁10bが内蔵されており、図示されるように環状に接続され冷媒回路を構成する。なお、以下の説明においては、例えば冷凍サイクル2a及び2bを総称するときには冷凍サイクル2と称するものとし、このことは他の機器においても同様とし、圧縮機3、四方弁4、空気熱交換器5、逆止弁6、過冷却熱交換器7、主膨張弁8、水熱交換器9、バイパス膨張弁10、ファン11とそれぞれ称するものとする。
Embodiment 1 FIG.
A first embodiment of the present invention is shown in FIG. FIG. 1 is a circuit diagram of a refrigerating and air-conditioning apparatus according to the present invention. A refrigeration cycle 2a, 2b having the same circuit configuration is mounted in a heat source unit 1 which is a refrigeration air conditioner. The refrigeration cycle 2a includes a compressor 3a, a four-way valve 4a, an air heat exchanger 5a that is a heat source side heat exchanger, check valves 6a to 6d, a supercooling heat exchanger 7a, a main expansion valve 8a that is a pressure reducing device, a load A water heat exchanger 9a, which is a side heat exchanger, and a bypass expansion valve 10a, which is a second pressure reducing device, are built in, and are connected in an annular shape as shown in the figure to constitute a refrigerant circuit. Similarly, in the refrigeration cycle 2b, a compressor 3b, a four-way valve 4b, an air heat exchanger 5b as a heat source side heat exchanger, check valves 6e to 6h, a supercooling heat exchanger 7b, and a pressure reducing device are used. A main expansion valve 8b, a water heat exchanger 9b that is a load-side heat exchanger, and a bypass expansion valve 10b that is a second pressure reducing device are incorporated, and are connected in a ring shape to form a refrigerant circuit as shown in the figure. . In the following description, for example, when the refrigeration cycles 2a and 2b are collectively referred to as the refrigeration cycle 2, the same applies to other devices, and the compressor 3, the four-way valve 4, and the air heat exchanger 5 are used. The check valve 6, the supercooling heat exchanger 7, the main expansion valve 8, the water heat exchanger 9, the bypass expansion valve 10, and the fan 11 are respectively referred to.
圧縮機3は、例えばDCブラシレスモータを搭載したスクロール圧縮機から構成されており、インバータ(図示せず)により回転数が制御され容量制御されるタイプのものである。空気熱交換器5は、プレートフィン熱交換器から構成されており、ファン11によって搬送される熱源機1周囲の空気と熱交換を行う。過冷却熱交換器7は、冷媒・冷媒熱交換器であり、プレート熱交換器から構成される。主膨張弁8及びバイパス膨張弁10は、開度が可変に制御される電子膨張弁から構成される。水熱交換器9は、プレート熱交換器から構成されており、熱負荷媒体である冷温水と冷媒との間で熱交換を行う。この冷凍空調装置の冷媒としては疑似共沸混合冷媒であるR410Aが用いられる。
冷媒回路は環状に接続され、水熱交換器9で冷水をつくる冷却運転では、圧縮機3、四方弁4、空気熱交換器5、逆止弁6a(6e)、過冷却熱交換器7の一方の流路、主膨張弁8、逆止弁6d(6h)、水熱交換器9、四方弁4、圧縮機3が環状に接続され、この順で冷媒が流れる。また過冷却熱交換器7を出た冷媒の一部が分岐され、バイパス膨張弁10、過冷却熱交換器7のもう一方の流路を経て圧縮機3の圧縮室にインジェクションされる。
水熱交換器9で温水をつくる加熱運転では、圧縮機3、四方弁4、水熱交換器9、逆止弁6b(6f)、過冷却熱交換器7の一方の流路、主膨張弁8、逆止弁6c(6g)、空気熱交換器5、四方弁4、圧縮機3が環状に接続され、この順で冷媒が流れる。また加熱運転においても過冷却熱交換器7を出た冷媒の一部が分岐され、上記冷却運転時と同様にバイパス膨張弁10、過冷却熱交換器7のもう一方の流路を経て圧縮機3の圧縮室にインジェクションされる。
このように冷却、加熱運転において過冷却熱交換器7から分岐後、バイパス膨張弁10、過冷却熱交換器7を経て圧縮機3にインジェクションされる回路にてエコノマイザ回路を構成する。
熱負荷媒体である冷温水は熱源機1の外部に設けられたポンプ12により搬送され、熱源機1内では点線の流路となり、冷凍サイクル2bの水熱交換器9b、冷凍サイクル2aの水熱交換器9aの順に流れる。水熱交換器9では、冷却運転時は冷媒と冷水が並行して流れる並行流となり、加熱運転時は冷媒と温水が対向して流れる対向流となるように流路構成されている。
The compressor 3 is composed of, for example, a scroll compressor equipped with a DC brushless motor, and is of a type in which the number of revolutions is controlled by an inverter (not shown) and the capacity is controlled. The air heat exchanger 5 is composed of a plate fin heat exchanger, and performs heat exchange with the air around the heat source unit 1 conveyed by the fan 11. The supercooling heat exchanger 7 is a refrigerant / refrigerant heat exchanger, and includes a plate heat exchanger. The main expansion valve 8 and the bypass expansion valve 10 are composed of electronic expansion valves whose opening degree is variably controlled. The water heat exchanger 9 is composed of a plate heat exchanger, and performs heat exchange between cold / hot water that is a heat load medium and a refrigerant. R410A, which is a pseudo azeotrope refrigerant, is used as the refrigerant of this refrigeration air conditioner.
The refrigerant circuit is annularly connected, and in the cooling operation in which cold water is generated by the water heat exchanger 9, the compressor 3, the four-way valve 4, the air heat exchanger 5, the check valve 6a (6e), and the supercooling heat exchanger 7 One flow path, the main expansion valve 8, the check valve 6d (6h), the water heat exchanger 9, the four-way valve 4, and the compressor 3 are connected in an annular shape, and the refrigerant flows in this order. Further, a part of the refrigerant exiting the supercooling heat exchanger 7 is branched and injected into the compression chamber of the compressor 3 through the bypass expansion valve 10 and the other flow path of the supercooling heat exchanger 7.
In the heating operation in which hot water is produced by the water heat exchanger 9, the compressor 3, the four-way valve 4, the water heat exchanger 9, the check valve 6b (6f), one flow path of the supercooling heat exchanger 7, the main expansion valve 8, the check valve 6c (6g), the air heat exchanger 5, the four-way valve 4, and the compressor 3 are connected in an annular shape, and the refrigerant flows in this order. In the heating operation, a part of the refrigerant exiting the supercooling heat exchanger 7 is branched, and the compressor is passed through the bypass expansion valve 10 and the other flow path of the supercooling heat exchanger 7 as in the cooling operation. 3 compression chambers.
In this way, an economizer circuit is constituted by a circuit that is branched from the supercooling heat exchanger 7 in the cooling and heating operation and then injected into the compressor 3 through the bypass expansion valve 10 and the supercooling heat exchanger 7.
Cold / warm water, which is a heat load medium, is conveyed by a pump 12 provided outside the heat source unit 1 and becomes a dotted channel in the heat source unit 1, and the water heat exchanger 9b of the refrigeration cycle 2b and the water heat of the refrigeration cycle 2a. It flows in the order of the exchanger 9a. The water heat exchanger 9 is configured to have a flow path so that the refrigerant and the cold water flow in parallel during the cooling operation, and the counter flow in which the refrigerant and hot water flow in the heating operation.
冷凍サイクル2a、2bには圧力センサ14a、14cが圧縮機3吸入側、圧力センサ14b、14dが圧縮機3吐出側に設けられており、それぞれ設置場所の冷媒圧力を計測する。また温度センサ15a、15iが圧縮機3吸入側、温度センサ15b、15jが圧縮機3吐出側、温度センサ15c、15kが空気熱交換器5の冷却運転時の出口側、温度センサ15d、15lが水熱交換器9の冷却運転時の入口側、温度センサ15e、15mがエコノマイザ回路上の過冷却熱交換器7流路の入口側、温度センサ15f、15nがエコノマイザ回路上の過冷却熱交換器7流路の出口側に設けられており、それぞれ設置場所の冷媒温度を計測する。また温度センサ15g、15oが水熱交換器9での冷温水の流入部、温度センサ15h、15pが水熱交換器9での冷温水の流出部に設けられており、それぞれ設置場所の冷温水の温度を計測する。温度センサ15sは熱源機1周囲の空気温度を計測するために設けられている。なお、以下の説明においては、圧力センサ14a〜14dを総称するときには圧力センサ14と称し、温度センサ15a〜15pを総称するときには温度センサ15と称するものとする。
計測制御装置13は、圧力センサ14、温度センサ15などの熱源機1の計測・運転情報や冷凍空調装置使用者から指示される運転内容に基づいて、圧縮機3の運転・停止や回転数、空気熱交換器5のファン送風量、主膨張弁8、バイパス膨張弁10の開度など各アクチュエータを制御する。
In the refrigeration cycles 2a and 2b, pressure sensors 14a and 14c are provided on the suction side of the compressor 3, and pressure sensors 14b and 14d are provided on the discharge side of the compressor 3, and the refrigerant pressure at the installation location is measured. Further, the temperature sensors 15a and 15i are the suction side of the compressor 3, the temperature sensors 15b and 15j are the discharge side of the compressor 3, the temperature sensors 15c and 15k are the outlet side during the cooling operation of the air heat exchanger 5, and the temperature sensors 15d and 15l are In the cooling operation of the water heat exchanger 9, the temperature sensors 15e and 15m are the inlet side of the supercooling heat exchanger 7 flow path on the economizer circuit, and the temperature sensors 15f and 15n are the supercooling heat exchanger on the economizer circuit. It is provided on the outlet side of the seven flow paths, and measures the refrigerant temperature at each installation location. Moreover, temperature sensors 15g and 15o are provided in the inflow part of the cold / hot water in the water heat exchanger 9, and temperature sensors 15h and 15p are provided in the outflow part of the cold / warm water in the water heat exchanger 9, respectively. Measure the temperature. The temperature sensor 15s is provided for measuring the air temperature around the heat source unit 1. In the following description, the pressure sensors 14a to 14d are collectively referred to as the pressure sensor 14, and the temperature sensors 15a to 15p are collectively referred to as the temperature sensor 15.
The measurement control device 13 operates / stops and rotates the compressor 3 based on the measurement / operation information of the heat source device 1 such as the pressure sensor 14 and the temperature sensor 15 and the operation content instructed by the user of the refrigeration air conditioner. Each actuator is controlled such as the fan air flow rate of the air heat exchanger 5, the opening of the main expansion valve 8, and the bypass expansion valve 10.
次に、この冷凍空調装置での運転動作について図1及び図2に基づいて説明する。図2は、この発明の実施の形態1における冷凍空調装置の圧力とエンタルピの関係を表した図であり、横軸はエンタルピを表し、縦軸は圧力を表している。冷凍サイクル2の運転動作は冷凍サイクル2a、2bとも同様となるので、代表して冷凍サイクル2aにおける動作を説明する。
まず冷却運転における冷媒回路の動作について説明する。冷却運転においては、四方弁4aの流路は図1の実線方向に設定される。圧縮機3aから吐出された高温高圧(Ph)のガス冷媒(図2点A)は、四方弁4aを経て空気熱交換器5aに流入し、凝縮器となる空気熱交換器5aで放熱しながら凝縮・液化する(図2点B)。空気熱交換器5aを出た高圧の液冷媒は逆止弁6aを経て、過冷却熱交換器7aで、エコノマイザ回路を流れる冷媒によりさらに冷却され(図2点C)、温度低下し主膨張弁8aに流入する。主膨張弁8aにて低圧(Pl)に減圧された二相状態の冷媒は(図2点D)、逆止弁6dを経て蒸発器となる水熱交換器9aにて、蒸発ガス化しながら吸熱し、負荷側熱媒体である水を冷却し冷水を生成する。水熱交換器9aを出た冷媒は、四方弁4aを経て圧縮機3aに吸入される(図2点E)。過冷却熱交換器7aを出た高圧の液冷媒の一部はエコノマイザ回路にバイパスされ、バイパス膨張弁10aにて、中間圧(Pm)まで減圧された後(図2点F)、過冷却熱交換器7aのもう一方の流路に流入し、空気熱交換器5aを出た高圧液冷媒と熱交換し加熱蒸発される(図2点G)。エコノマイザ回路を流れる冷媒は、その後圧縮機3a内の圧縮途中の圧縮室にインジェクションされ、吸入状態(図2点E)から圧縮された冷媒(図2点H)と混合した後(図2点I)、高圧(Ph)まで圧縮され、高温高圧のガス冷媒(図2点A)となる。
次に冷却運転における冷水の動作について説明する。冷水はポンプ12によって駆動される。低温の、例えば7℃の冷水はファンコイルなど負荷側装置に流入し、そこで負荷側装置周囲に冷熱を供給しながら冷水そのものの温度は上昇し、例えば12℃まで上昇した後で、熱源機1に流入する。熱源機1に流入した冷水は冷凍サイクル2bの水熱交換器9bにて冷媒により冷却され温度低下し、例えば9.5℃となって流出し、次いで冷凍サイクル2aの水熱交換器9aに流入する。ここで冷水は冷媒により冷却され、さらに温度低下し、例えば7℃となって、水熱交換器9aを流出し、熱源機1を流出する。その後冷水は再び負荷側装置に流入する。
Next, the operation | movement operation | movement in this refrigeration air conditioning apparatus is demonstrated based on FIG.1 and FIG.2. FIG. 2 is a diagram showing the relationship between the pressure and enthalpy of the refrigerating and air-conditioning apparatus according to Embodiment 1 of the present invention, in which the horizontal axis represents enthalpy and the vertical axis represents pressure. Since the operation of the refrigeration cycle 2 is the same as that of the refrigeration cycles 2a and 2b, the operation in the refrigeration cycle 2a will be described as a representative.
First, the operation of the refrigerant circuit in the cooling operation will be described. In the cooling operation, the flow path of the four-way valve 4a is set in the direction of the solid line in FIG. The high-temperature and high-pressure (Ph) gas refrigerant (point A in FIG. 2) discharged from the compressor 3a flows into the air heat exchanger 5a through the four-way valve 4a and dissipates heat in the air heat exchanger 5a serving as a condenser. Condensation and liquefaction (point B in FIG. 2). The high-pressure liquid refrigerant exiting the air heat exchanger 5a passes through the check valve 6a, and is further cooled by the refrigerant flowing through the economizer circuit in the supercooling heat exchanger 7a (point C in FIG. 2). It flows into 8a. The refrigerant in the two-phase state decompressed to a low pressure (Pl) by the main expansion valve 8a (D in FIG. 2) absorbs heat while evaporating and gasifying in the water heat exchanger 9a that becomes the evaporator through the check valve 6d. Then, water that is the load-side heat medium is cooled to generate cold water. The refrigerant exiting the water heat exchanger 9a is sucked into the compressor 3a through the four-way valve 4a (point E in FIG. 2). A part of the high-pressure liquid refrigerant exiting the supercooling heat exchanger 7a is bypassed to the economizer circuit, and is depressurized to the intermediate pressure (Pm) by the bypass expansion valve 10a (point F in FIG. 2). The refrigerant flows into the other flow path of the exchanger 7a, exchanges heat with the high-pressure liquid refrigerant discharged from the air heat exchanger 5a, and is evaporated by heating (point G in FIG. 2). The refrigerant flowing through the economizer circuit is then injected into the compression chamber in the compressor 3a in the middle of compression and mixed with the refrigerant (point H in FIG. 2) from the suction state (point E in FIG. 2) (point I in FIG. 2). ), Compressed to a high pressure (Ph), and becomes a high-temperature and high-pressure gas refrigerant (point A in FIG. 2).
Next, the operation of cold water in the cooling operation will be described. The cold water is driven by the pump 12. Low temperature, for example, 7 ° C. chilled water flows into a load side device such as a fan coil, where the temperature of the chilled water itself rises while supplying cold heat around the load side device. Flow into. The cold water that has flowed into the heat source unit 1 is cooled by the refrigerant in the water heat exchanger 9b of the refrigeration cycle 2b and drops in temperature to, for example, 9.5 ° C., and then flows out to the water heat exchanger 9a of the refrigeration cycle 2a. To do. Here, the cold water is cooled by the refrigerant, and further drops in temperature, for example, 7 ° C., flows out of the water heat exchanger 9a, and flows out of the heat source unit 1. Thereafter, the cold water flows into the load side device again.
次に加熱運転における冷媒回路の動作について説明する。加熱運転においても、冷凍サイクル2a、2bの動作は同様となるので、代表して冷凍サイクル2aにおける動作を説明する。加熱運転では四方弁4aの流路は図1の点線方向に設定される。加熱運転における冷媒の状態変化も冷却運転とほぼ同様であり、図2に示される状態変化となる。圧縮機3aから吐出された高温高圧(Ph)のガス冷媒(図2点A)は、四方弁4aを経て水熱交換器9aに流入し、凝縮器となる水熱交換器9aで放熱しながら凝縮・液化する(図2点B)。この際、液媒体である水を加熱し温水を生成する。水熱交換器9aを出た高圧の液冷媒は逆止弁6bを経て、過冷却熱交換器7aで、エコノマイザ回路を流れる冷媒によりさらに冷却され(図2点C)、温度低下し主膨張弁8aに流入する。主膨張弁8aにて低圧(Pl)に減圧され二相状態の冷媒となり(図2点D)、逆止弁6cを経て蒸発器となる空気熱交換器5aに流入し、空気熱交換器5aにて、蒸発ガス化され、四方弁4aを経て圧縮機3aに吸入される(図2点E)。過冷却熱交換器7aを出た高圧の液冷媒の一部はエコノマイザ回路にバイパスされ、バイパス膨張弁10aにて、中間圧(Pm)まで減圧された後、過冷却熱交換器7aのもう一方の流路に流入し、水熱交換器9aを出た高圧液冷媒と熱交換し加熱蒸発される(図2点G)。エコノマイザ回路を流れる冷媒は、その後圧縮機3a内の圧縮途中の圧縮室にインジェクションされ、吸入状態(図2点E)から圧縮された冷媒(図2点H)と混合した後(図2点I)、高圧(Ph)まで圧縮され、高温高圧のガス冷媒(図2点A)となる。
次に加熱運転における温水の動作について説明する。温水はポンプ12によって駆動される。高温の、例えば45℃の温水はファンコイルなど負荷側装置に流入し、そこで負荷側装置周囲に温熱を供給しながら温水そのものの温度は低下し、例えば40℃まで低下した後で、熱源機1に流入する。熱源機1に流入した温水は冷凍サイクル2bの水熱交換器9bにて冷媒により加熱され温度上昇し、例えば42.5℃となって流出し、次いで冷凍サイクル2aの水熱交換器9aに流入する。ここで温水は冷媒により加熱され、さらに温度上昇し、例えば45℃となって、水熱交換器9aを流出し、熱源機1を流出する。その後温水は再び負荷側装置に流入する。
Next, the operation of the refrigerant circuit in the heating operation will be described. Since the operation of the refrigeration cycles 2a and 2b is the same in the heating operation, the operation in the refrigeration cycle 2a will be described as a representative. In the heating operation, the flow path of the four-way valve 4a is set in the direction of the dotted line in FIG. The state change of the refrigerant in the heating operation is almost the same as that in the cooling operation, and the state change shown in FIG. The high-temperature and high-pressure (Ph) gas refrigerant (point A in FIG. 2) discharged from the compressor 3a flows into the hydrothermal exchanger 9a via the four-way valve 4a, and dissipates heat in the hydrothermal exchanger 9a serving as a condenser. It condenses and liquefies (Fig. 2, point B). At this time, water as a liquid medium is heated to generate hot water. The high-pressure liquid refrigerant exiting the water heat exchanger 9a passes through the check valve 6b, and is further cooled by the refrigerant flowing through the economizer circuit in the supercooling heat exchanger 7a (point C in FIG. 2). It flows into 8a. The refrigerant is reduced to a low pressure (Pl) by the main expansion valve 8a to be a two-phase refrigerant (point D in FIG. 2), flows into the air heat exchanger 5a serving as an evaporator through the check valve 6c, and then flows into the air heat exchanger 5a. The gas is evaporated and is sucked into the compressor 3a through the four-way valve 4a (point E in FIG. 2). A part of the high-pressure liquid refrigerant exiting the supercooling heat exchanger 7a is bypassed to the economizer circuit, and after being depressurized to an intermediate pressure (Pm) by the bypass expansion valve 10a, the other side of the supercooling heat exchanger 7a. The heat is exchanged with the high-pressure liquid refrigerant exiting the water heat exchanger 9a and evaporated by heating (point G in FIG. 2). The refrigerant flowing through the economizer circuit is then injected into the compression chamber in the compressor 3a in the middle of compression and mixed with the refrigerant (point H in FIG. 2) from the suction state (point E in FIG. 2) (point I in FIG. 2). ), Compressed to a high pressure (Ph), and becomes a high-temperature and high-pressure gas refrigerant (point A in FIG. 2).
Next, the operation of hot water in the heating operation will be described. The hot water is driven by the pump 12. High-temperature hot water, for example, 45 ° C. flows into a load-side device such as a fan coil, where the temperature of the hot water itself decreases while supplying heat around the load-side device. Flow into. The hot water flowing into the heat source unit 1 is heated by the refrigerant in the water heat exchanger 9b of the refrigeration cycle 2b and rises in temperature, for example, flows out to 42.5 ° C., and then flows into the water heat exchanger 9a of the refrigeration cycle 2a. To do. Here, the hot water is heated by the refrigerant and further rises in temperature, for example, 45 ° C., flows out of the water heat exchanger 9a, and flows out of the heat source unit 1. Thereafter, the hot water flows into the load side device again.
次に、この冷凍空調装置での制御動作について説明する。始めに冷却運転について図3に基づいて説明する。まず冷凍空調装置使用者などにより、負荷側装置に供給される冷水の目標温度が設定される(S11)。また負荷側装置の運転状況に応じて冷水を送水するポンプ12の流量が変更される(S12)。   Next, the control operation in this refrigeration air conditioner will be described. First, the cooling operation will be described with reference to FIG. First, a target temperature of cold water supplied to the load side device is set by a user of the refrigeration air conditioner or the like (S11). Moreover, the flow volume of the pump 12 which supplies cold water according to the driving | running state of a load side apparatus is changed (S12).
熱源機1内では、温度センサ15oで検知される冷水の流入温度、および負荷側装置に供給する冷水の目標温度に応じて、冷凍サイクル2a、2bの水熱交換器9出口の冷水温度目標値が設定される(S13)。この冷水温度目標値は、水熱交換器9出入口の冷水温度差が冷凍サイクル2a、2bの熱交換器の定格熱交換量を示す容量、もしくは水熱交換器9の伝熱面積など定格熱交換量を決定する値に対して、比例するように決定される。例えば、熱源機1に流入する冷水温度が12℃、供給される冷水の目標温度が7℃、冷凍サイクル2a、2bの水熱交換器9の定格容量が等しい場合には、熱源機1全体の冷水温度差が12−7=5℃となるため、冷凍サイクル2a、2bともに冷水温度差が5×1/2=2.5℃となるように設定される。従って、水熱交換器9が冷水流路上流に配置される冷凍サイクル2bでは、水熱交換器9b出口の冷水温度目標値は12−2.5=9.5℃に設定される、水熱交換器9が冷水流路下流に配置される冷凍サイクル2aでは、水熱交換器9a出口の冷水温度目標値は9.5−2.5=7℃に設定され、これは当然負荷側装置に供給する冷水温度の目標値と等しくなる。   In the heat source unit 1, the target value of the cold water temperature at the outlet of the water heat exchanger 9 of the refrigeration cycles 2a and 2b according to the inflow temperature of the cold water detected by the temperature sensor 15o and the target temperature of the cold water supplied to the load side device. Is set (S13). This chilled water temperature target value is the rated heat exchange such as the capacity at which the chilled water temperature difference at the inlet / outlet of the water heat exchanger 9 indicates the rated heat exchange amount of the heat exchanger of the refrigeration cycle 2a, 2b, or the heat transfer area of the water heat exchanger 9. It is determined to be proportional to the value that determines the quantity. For example, when the temperature of the cold water flowing into the heat source unit 1 is 12 ° C., the target temperature of the supplied cold water is 7 ° C., and the rated capacities of the water heat exchangers 9 of the refrigeration cycles 2a and 2b are equal, Since the cold water temperature difference is 12−7 = 5 ° C., the refrigeration cycles 2a and 2b are set so that the cold water temperature difference is 5 × 1/2 = 2.5 ° C. Therefore, in the refrigeration cycle 2b in which the water heat exchanger 9 is arranged upstream of the cold water flow path, the target value of the cold water temperature at the outlet of the water heat exchanger 9b is set to 12−2.5 = 9.5 ° C., In the refrigeration cycle 2a in which the exchanger 9 is arranged downstream of the chilled water flow path, the target value of the chilled water temperature at the outlet of the water heat exchanger 9a is set to 9.5-2.5 = 7 ° C. It becomes equal to the target value of the supplied cold water temperature.
次に、各冷凍サイクル2a、2bの運転制御について説明する。ここで制御方法は、各冷凍サイクル共通となるので、冷凍サイクル2aについて説明する。まず、圧縮機3aの回転数、空気熱交換器5aへの送風量、主膨張弁8aの開度、バイパス膨張弁10aの開度を初期値に設定して運転を行う(S14)。空気熱交換器5aの送風量の初期設定値は温度センサ15sで検知される外気温度、およびあらかじめ計測制御装置13に記憶された所定値とを比較して決定される。ここで外気温度と比較する所定値は圧縮機の運転容量、熱交換器性能など機器性能に基づいて定められ、冷凍サイクルの高圧(圧縮機3a吐出冷媒の圧力)が低下しすぎないようにするため、外気温度が高い場合は高風量、低い場合は低風量に設定される。   Next, operation control of each refrigeration cycle 2a, 2b will be described. Here, since the control method is common to each refrigeration cycle, the refrigeration cycle 2a will be described. First, operation is performed with the rotation speed of the compressor 3a, the amount of air blown to the air heat exchanger 5a, the opening of the main expansion valve 8a, and the opening of the bypass expansion valve 10a being set to initial values (S14). The initial setting value of the air flow rate of the air heat exchanger 5a is determined by comparing the outside air temperature detected by the temperature sensor 15s and a predetermined value stored in the measurement control device 13 in advance. Here, the predetermined value to be compared with the outside air temperature is determined based on equipment performance such as the operating capacity of the compressor and heat exchanger performance, so that the high pressure of the refrigeration cycle (pressure of the refrigerant discharged from the compressor 3a) does not decrease too much. Therefore, the high air volume is set when the outside air temperature is high, and the low air volume is set when it is low.
そして、この状態で運転した後、装置運転状態に応じて各アクチュエータを制御する。まず圧縮機3の回転数は、温度センサ15hで検知される水熱交換器9出口の冷水温度が前述した目標値となるように制御される。圧縮機3の回転数が高いと、冷媒流量が増加するため冷却能力が増加し、水がより冷却されるため、水熱交換器9出口の水温は低下する。逆に、圧縮機3の回転数が低いと、水熱交換器9出口の水温は上昇する。そこで水熱交換器9出口の水温と目標値とを比較し(S15)、水温が高い場合は圧縮機3の回転数を増加させ、水温が低い場合は圧縮機3の回転数を減少させる(S16)。   And after driving | running in this state, each actuator is controlled according to an apparatus operating state. First, the rotation speed of the compressor 3 is controlled so that the cold water temperature at the outlet of the water heat exchanger 9 detected by the temperature sensor 15h becomes the above-described target value. If the rotation speed of the compressor 3 is high, the refrigerant flow rate increases, the cooling capacity increases, and the water is further cooled, so the water temperature at the outlet of the water heat exchanger 9 decreases. Conversely, when the rotation speed of the compressor 3 is low, the water temperature at the outlet of the water heat exchanger 9 rises. Therefore, the water temperature at the outlet of the water heat exchanger 9 is compared with the target value (S15). When the water temperature is high, the rotation speed of the compressor 3 is increased, and when the water temperature is low, the rotation speed of the compressor 3 is decreased ( S16).
次に、空気熱交換器5の送風量であるが、この送風量は基本的に初期設定値にて運転を行う。ただし、運転条件によって、圧力センサ14bで検知される高圧(圧力)が所定範囲内からはずれるような場合には、高圧が所定範囲内であるかどうかを確認し(S17)、高圧が過度に上昇した場合は圧縮機3a保護のために風量を増加させる制御を行う。また、高圧が過度に低下した場合は、主膨張弁8の開度制御を行っても低圧(圧縮機3a吸入冷媒の圧力)が大きく低下し、冷媒蒸発温度が氷点下以下に低下し、冷水が凍結する恐れが出てくるので、高圧の過度の低下を抑制するように風量を減少させる制御を行う(S18)。   Next, although it is the air flow rate of the air heat exchanger 5, this air flow rate is basically operated at an initial set value. However, if the high pressure (pressure) detected by the pressure sensor 14b deviates from the predetermined range depending on the operating conditions, it is confirmed whether the high pressure is within the predetermined range (S17) and the high pressure rises excessively. In such a case, control for increasing the air volume is performed to protect the compressor 3a. If the high pressure is excessively reduced, the low pressure (pressure of the refrigerant sucked by the compressor 3a) is greatly reduced even if the opening degree of the main expansion valve 8 is controlled, the refrigerant evaporation temperature is lowered below the freezing point, Since there is a risk of freezing, control is performed to reduce the air volume so as to suppress an excessive decrease in high pressure (S18).
次に、主膨張弁8aの開度であるが、蒸発器となる水熱交換器9aの出口であり、圧縮機3a吸入の状態(図2点E)の冷媒過熱度SHを演算し(S19)、この冷媒過熱度SHが、予め設定された目標値、例えば1℃となるように制御される。ここで水熱交換器9aの出口であり圧縮機3a吸入の冷媒過熱度SHは、(温度センサ15a検知温度(圧縮機3の吸入温度))−(圧力センサ14aから換算される冷媒飽和温度)で演算される値を用いる。
主膨張弁8aの開度が小さくなると、水熱交換器9aを流れる冷媒流量は減少し、水熱交換器9a出口の冷媒過熱度SHは大きくなり、逆に主膨張弁8aの開度を大きくすると水熱交換器9aの冷媒過熱度SHは小さくなる。そこで、圧縮機3a吸入(水熱交換器9a出口)の冷媒過熱度SHと目標値とを比較し(S20)、冷媒過熱度SHが目標値より大きい場合には、主膨張弁8aの開度を大きく制御し、冷媒過熱度SHが目標値より小さい場合には主膨張弁8aの開度を小さく制御する(S21)。
Next, the opening degree of the main expansion valve 8a, which is the outlet of the water heat exchanger 9a serving as an evaporator, calculates the refrigerant superheat degree SH in the suction state of the compressor 3a (point E in FIG. 2) (S19). ), The refrigerant superheat degree SH is controlled to be a preset target value, for example, 1 ° C. Here, the refrigerant superheat degree SH at the outlet of the water heat exchanger 9a and sucked into the compressor 3a is (temperature sensor 15a detected temperature (intake temperature of the compressor 3)) − (refrigerant saturation temperature converted from the pressure sensor 14a). Use the value calculated in.
When the opening of the main expansion valve 8a decreases, the flow rate of the refrigerant flowing through the water heat exchanger 9a decreases, the refrigerant superheat degree SH at the outlet of the water heat exchanger 9a increases, and conversely, the opening of the main expansion valve 8a increases. Then, the refrigerant superheat degree SH of the water heat exchanger 9a becomes small. Accordingly, the refrigerant superheat degree SH of the compressor 3a suction (water heat exchanger 9a outlet) is compared with the target value (S20), and when the refrigerant superheat degree SH is larger than the target value, the opening degree of the main expansion valve 8a When the refrigerant superheat degree SH is smaller than the target value, the opening degree of the main expansion valve 8a is controlled to be small (S21).
次に、バイパス膨張弁10aの開度であるが、エコノマイザ回路上の過冷却熱交換器7a出口(図2点G)の冷媒過熱度SHecoを演算し(S22)、この冷媒過熱度SHecoが、予め設定された目標値、例えば2℃となるように制御される。ここで過冷却熱交換器7出口の冷媒過熱度SHecoは、温度センサ15f検知温度−温度センサ15e検知温度で演算される値を用いる。
バイパス膨張弁10aの開度が小さくなると、エコノマイザ回路を流れる冷媒流量は減少し、エコノマイザ回路上の過冷却熱交換器7a出口の冷媒過熱度SHは大きくなり、逆にバイパス膨張弁10aの開度を大きくすると過冷却熱交換器7a出口の冷媒過熱度SHは小さくなる。そこで、過冷却熱交換器7a出口の冷媒過熱度SHecoと目標値とを比較し(S23)、冷媒過熱度SHecoが目標値より大きい場合には、バイパス膨張弁10aの開度を大きく制御し、冷媒過熱度SHecoが目標値より小さい場合にはバイパス膨張弁10aの開度を小さく制御する(S24)。
Next, the opening degree of the bypass expansion valve 10a, the refrigerant superheat degree Sheco at the outlet of the supercooling heat exchanger 7a on the economizer circuit (point G in FIG. 2) is calculated (S22), and the refrigerant superheat degree Sheco is It is controlled so as to be a preset target value, for example, 2 ° C. Here, the refrigerant superheat degree SHeco at the outlet of the supercooling heat exchanger 7 uses a value calculated from the temperature sensor 15f detected temperature-temperature sensor 15e detected temperature.
When the opening degree of the bypass expansion valve 10a is reduced, the refrigerant flow rate flowing through the economizer circuit is decreased, the refrigerant superheat degree SH at the outlet of the supercooling heat exchanger 7a on the economizer circuit is increased, and conversely, the opening degree of the bypass expansion valve 10a. Is increased, the refrigerant superheat degree SH at the outlet of the supercooling heat exchanger 7a is decreased. Therefore, the refrigerant superheat degree Sheco at the outlet of the supercooling heat exchanger 7a is compared with the target value (S23), and if the refrigerant superheat degree Sheco is larger than the target value, the opening degree of the bypass expansion valve 10a is largely controlled, When the refrigerant superheat degree Sheco is smaller than the target value, the opening degree of the bypass expansion valve 10a is controlled to be small (S24).
次に、負荷状況に応じて、負荷側装置の運転状況が変化し、冷水を送水するポンプ12の流量が変更される(S25)が、ポンプ流量が増減した場合の制御動作について説明する。まず、冷凍サイクル2a、2bともに動作している状況でポンプ流量、すなわち冷水流量が増加した場合の運転動作について説明する。この場合、冷水流量が増加する一方で、冷凍サイクル2a、2bの蒸発器熱交換量、即ち冷却能力は変化しないので、水熱交換器9出入口の冷水温度差が減少し、水熱交換器9出口の冷水温度が上昇する。そこで冷水出口温度が目標値となるように、各冷凍サイクル2の圧縮機3の回転数が増速される運転が行われる。   Next, the operation state of the load side device changes according to the load state, and the flow rate of the pump 12 that supplies cold water is changed (S25), but the control operation when the pump flow rate increases or decreases will be described. First, an operation operation when the pump flow rate, that is, the chilled water flow rate is increased while both the refrigeration cycles 2a and 2b are operating will be described. In this case, while the flow rate of the chilled water is increased, the heat exchange amount of the evaporators of the refrigeration cycles 2a and 2b, that is, the cooling capacity is not changed, so that the chilled water temperature difference at the inlet / outlet of the water heat exchanger 9 is reduced. The cold water temperature at the outlet rises. Therefore, an operation is performed in which the rotation speed of the compressor 3 of each refrigeration cycle 2 is increased so that the cold water outlet temperature becomes the target value.
逆に冷水流量が減少した場合の運転動作は以下のようになる。この場合、冷水流量が減少する一方で、冷凍サイクル2a、2bの蒸発器熱交換量、即ち冷却能力は変化しないので、水熱交換器9出入口の冷水温度差が増加し、水熱交換器9出口の冷水温度が低下する。そこで冷水出口温度が目標値となるように、各冷凍サイクル2の圧縮機3の回転数が減速される運転が行われる。   Conversely, the operation when the cold water flow rate decreases is as follows. In this case, while the flow rate of the chilled water is decreased, the heat exchange amount of the evaporators of the refrigeration cycles 2a and 2b, that is, the cooling capacity is not changed, so the chilled water temperature difference at the inlet / outlet of the water heat exchanger 9 is increased, and the water heat exchanger 9 The cold water temperature at the outlet decreases. Therefore, an operation is performed in which the rotational speed of the compressor 3 of each refrigeration cycle 2 is decelerated so that the cold water outlet temperature becomes the target value.
圧縮機3の回転数については、動作下限値があるので、冷水流量が、ある程度以上低下し、圧縮機3回転数が動作下限値に近づき、所定の回転数以下となったら、一方の冷凍サイクル2の圧縮機を停止し、運転させる冷凍サイクル2の数を減少させる。例えば、圧縮機3の下限回転数が20rpsの場合には、圧縮機3の回転数が30rpsとなった段階で、圧縮機の停止動作を行わせる。このとき、冷水上流流路に接続される冷凍サイクルから順に停止動作を行う。冷凍サイクル2の停止に伴って、水熱交換器9出口の冷水温度目標値が再設定される。この実施の形態では、冷凍サイクルの数が2から1に減少され、冷水流路の最も下流側に位置する冷凍サイクル2aのみの運転となるので、この場合の冷水出口温度目標値は、冷凍サイクルの数によらず7℃のままとなる。そして上述した冷凍サイクルの運転制御が実施される。   Since there is an operation lower limit value for the rotation speed of the compressor 3, when the chilled water flow rate decreases to a certain degree or more and the rotation speed of the compressor 3 approaches the operation lower limit value and becomes a predetermined rotation speed or less, one refrigeration cycle 2 compressors are stopped and the number of refrigeration cycles 2 to be operated is reduced. For example, when the lower limit rotation speed of the compressor 3 is 20 rps, the compressor is stopped when the rotation speed of the compressor 3 reaches 30 rps. At this time, the stop operation is sequentially performed from the refrigeration cycle connected to the cold water upstream flow path. As the refrigeration cycle 2 is stopped, the target value of the cold water temperature at the outlet of the water heat exchanger 9 is reset. In this embodiment, the number of refrigeration cycles is reduced from 2 to 1, and only the refrigeration cycle 2a located on the most downstream side of the chilled water flow path is operated, so the chilled water outlet temperature target value in this case is Regardless of the number, the temperature remains at 7 ° C. Then, the above-described operation control of the refrigeration cycle is performed.
冷凍サイクル2aのみ動作している状況で冷水流量が減少した場合には、冷水出口温度が目標値となるように、冷凍サイクル2aの圧縮機3aの回転数が減速される運転が行われる。さらに冷水流量が減少し、圧縮機3の回転数が下限値となるまで制御された場合には、下限回転数での運転を継続する。さらに冷水流量が減少し、冷水出口温度が目標値より所定温度、例えば1℃以上下回り、6℃以下となった場合には、冷凍サイクル2aの圧縮機3も停止動作を行い、熱源機1の運転を停止する。   When the chilled water flow rate is reduced while only the refrigeration cycle 2a is operating, an operation is performed in which the rotational speed of the compressor 3a of the refrigeration cycle 2a is reduced so that the cold water outlet temperature becomes the target value. Further, when the flow rate of the cold water is decreased and the rotation speed of the compressor 3 is controlled until it reaches the lower limit value, the operation at the lower limit rotation speed is continued. When the chilled water flow rate further decreases and the chilled water outlet temperature falls below a target temperature by a predetermined temperature, for example, 1 ° C. or more and becomes 6 ° C. or less, the compressor 3 of the refrigeration cycle 2a also stops and Stop operation.
熱源機1が停止している状況で、温度センサ15hで検知される熱源機1の冷水出口温度が負荷側への供給目標温度より、所定温度、例えば1℃以上上回り、8℃以上となった場合には、熱源機1の運転を行う。そして、冷水流路の下流に配置される冷凍サイクル2aの圧縮機3aを起動し、水熱交換器9a出口の冷水温度が目標値となるように圧縮機3aの回転数制御を行う。   While the heat source unit 1 is stopped, the cold water outlet temperature of the heat source unit 1 detected by the temperature sensor 15h exceeds the target supply temperature to the load side by a predetermined temperature, for example, 1 ° C. or more, and becomes 8 ° C. or more. In that case, the heat source machine 1 is operated. And the compressor 3a of the refrigerating cycle 2a arrange | positioned downstream of a cold-water flow path is started, and the rotation speed control of the compressor 3a is performed so that the cold-water temperature of the water heat exchanger 9a exit may become a target value.
冷凍サイクル2aのみ動作している状況で冷水流量が増加した場合には、冷水出口温度が目標値となるように、冷凍サイクル2aの圧縮機3aの回転数が増速される運転が行われる。さらに冷水流量が増加し、圧縮機3aの回転数が所定値以上となるまで制御された場合には、冷水流路の上流に配置される冷凍サイクル2bの圧縮機3bを起動する。冷凍サイクル2bの運転に伴って、水熱交換器9出口の冷水温度目標値が再設定され、各冷凍サイクルで水熱交換器9出口の冷水温度が目標値となるように圧縮機の回転数制御が実施される。   When the chilled water flow rate increases while only the refrigeration cycle 2a is operating, an operation is performed in which the rotational speed of the compressor 3a of the refrigeration cycle 2a is increased so that the chilled water outlet temperature becomes the target value. When the flow rate of the chilled water further increases and the rotation speed of the compressor 3a is controlled to a predetermined value or more, the compressor 3b of the refrigeration cycle 2b disposed upstream of the chilled water flow path is started. With the operation of the refrigeration cycle 2b, the chilled water temperature target value at the outlet of the water heat exchanger 9 is reset, and the rotation speed of the compressor is set so that the chilled water temperature at the outlet of the water heat exchanger 9 becomes the target value in each refrigeration cycle. Control is implemented.
次に、この冷凍空調装置での加熱運転の制御動作について図4に基づいて説明する。まず冷凍空調装置使用者などにより、負荷側装置に供給される温水の目標温度が設定される(S31)。また負荷側装置の運転状況に応じて温水を送水するポンプ12の流量が変更される(S32)。   Next, the control operation of the heating operation in this refrigeration air conditioner will be described based on FIG. First, a target temperature of hot water supplied to the load side device is set by a user of the refrigeration air conditioner or the like (S31). Moreover, the flow volume of the pump 12 which supplies warm water according to the driving | running state of a load side apparatus is changed (S32).
熱源機1内では、温度センサ15oで検知される温水の流入温度、および負荷側装置に供給する温水の目標温度に応じて、冷凍サイクル2a、2bの水熱交換器9出口の温水温度目標値が設定される(S33)。この温水温度目標値は、水熱交換器9出入口の温水温度差が冷凍サイクル2a、2bの熱交換器の定格熱交換量を示す容量、もしくは水熱交換器9の伝熱面積など定格熱交換量を決定する値に対して、比例するように決定される。例えば、熱源機1に流入する温水温度が40℃、供給される温水の目標温度が45℃、冷凍サイクル2a、2bの水熱交換器9の定格容量が等しい場合には、熱源機1全体の温水温度差が45−40=5℃となるため、冷凍サイクル2a、2bともに温水温度差が5×1/2=2.5℃となるように設定される。従って、水熱交換器9が温水流路上流に配置される冷凍サイクル2bでは、水熱交換器9b出口の温水温度目標値は40+2.5=42.5℃に設定される、水熱交換器9が温水流路下流に配置される冷凍サイクル2aでは、水熱交換器9a出口の温水温度目標値は42.5+2.5=45℃に設定され、これは当然負荷側装置に供給する温水温度の目標値と等しくなる。   In the heat source unit 1, the hot water temperature target value at the outlet of the water heat exchanger 9 of the refrigeration cycle 2a, 2b according to the inflow temperature of the hot water detected by the temperature sensor 15o and the target temperature of the hot water supplied to the load side device. Is set (S33). This hot water temperature target value is a rated heat exchange such as the capacity at which the temperature difference of the hot water at the inlet / outlet of the water heat exchanger 9 indicates the rated heat exchange amount of the heat exchanger of the refrigeration cycle 2a, 2b, or the heat transfer area of the water heat exchanger 9. It is determined to be proportional to the value that determines the quantity. For example, when the temperature of hot water flowing into the heat source unit 1 is 40 ° C., the target temperature of the supplied hot water is 45 ° C., and the rated capacities of the water heat exchangers 9 of the refrigeration cycles 2a and 2b are equal, Since the hot water temperature difference is 45−40 = 5 ° C., the refrigeration cycles 2 a and 2 b are set so that the hot water temperature difference is 5 × 1/2 = 2.5 ° C. Accordingly, in the refrigeration cycle 2b in which the water heat exchanger 9 is arranged upstream of the hot water flow path, the target value of the hot water temperature at the outlet of the water heat exchanger 9b is set to 40 + 2.5 = 42.5 ° C. In the refrigeration cycle 2a in which 9 is disposed downstream of the hot water flow path, the target value of the hot water temperature at the outlet of the water heat exchanger 9a is set to 42.5 + 2.5 = 45 ° C. This is naturally the hot water temperature supplied to the load side device. Equal to the target value of.
次に、各冷凍サイクル2a、2bの運転制御について説明する。ここで制御方法は、各冷凍サイクル共通となるので、冷凍サイクル2aについて説明する。
まず、圧縮機3aの回転数、空気熱交換器5aの送風量、主膨張弁8aの開度、バイパス膨張弁10aの開度を初期値に設定して運転を行う(S34)。ここで空気熱交換器5送風量の初期設定値は温度センサ15sで検知される外気温度およびあらかじめ計測制御装置13に記憶された所定値とを比較して決定され、外気温度が低い場合は高風量、高い場合は低風量に設定される。
Next, operation control of each refrigeration cycle 2a, 2b will be described. Here, since the control method is common to each refrigeration cycle, the refrigeration cycle 2a will be described.
First, the operation is performed with the rotation speed of the compressor 3a, the amount of air blown by the air heat exchanger 5a, the opening of the main expansion valve 8a, and the opening of the bypass expansion valve 10a being set to initial values (S34). Here, the initial setting value of the air heat exchanger 5 is determined by comparing the outside air temperature detected by the temperature sensor 15s with a predetermined value stored in advance in the measurement control device 13, and is high when the outside air temperature is low. If the air volume is high, the air volume is set low.
そして、この状態で運転した後、装置運転状態に応じて各アクチュエータを制御する。まず圧縮機3の回転数は、温度センサ15hで検知される水熱交換器9出口の温水温度が予め設定された目標値となるように制御される。圧縮機3の回転数が高いと、冷媒流量が増加するため装置の冷却能力が増加し、水がより加熱されるため、水熱交換器9出口の水温は上昇する。逆に、圧縮機3の回転数が低いと、水熱交換器9出口の水温は低下する。そこで水熱交換器9出口の水温と目標値とを比較し(S35)、水温が低い場合は圧縮機3の回転数を増加させ、水温が高い場合は圧縮機3の回転数を減少させる(S36)。   And after driving | running in this state, each actuator is controlled according to an apparatus operating state. First, the rotation speed of the compressor 3 is controlled so that the hot water temperature at the outlet of the water heat exchanger 9 detected by the temperature sensor 15h becomes a preset target value. If the rotation speed of the compressor 3 is high, the refrigerant flow rate increases, the cooling capacity of the apparatus increases, and the water is further heated, so the water temperature at the outlet of the water heat exchanger 9 rises. On the contrary, when the rotation speed of the compressor 3 is low, the water temperature at the outlet of the water heat exchanger 9 is lowered. Therefore, the water temperature at the outlet of the water heat exchanger 9 is compared with the target value (S35). When the water temperature is low, the rotational speed of the compressor 3 is increased, and when the water temperature is high, the rotational speed of the compressor 3 is decreased ( S36).
次に、空気熱交換器5の送風量であるが、この送風量は基本的に初期設定値にて運転を行う。状況として高外気温(たとえば15℃くらい)に、加熱運転を行った場合に、圧縮機の負荷が過大となるのを防止するため風量を低下させ、冷凍サイクルの低圧を低下し、圧縮機の搬送流量を低下することで、圧縮機駆動の負荷を低減する場合があるが、この実施の形態が対象とする冷凍空調装置が用いられるビル用空調などの場合、高外気温時に暖房負荷が発生することはほとんどないため、上記の通り初期設定値にて運転を行う。   Next, although it is the air flow rate of the air heat exchanger 5, this air flow rate is basically operated at an initial set value. As a situation, when heating operation is performed at a high outside air temperature (for example, about 15 ° C.), the air flow is reduced to prevent the compressor load from becoming excessive, the low pressure of the refrigeration cycle is reduced, and the compressor Reducing the transport flow rate may reduce the compressor drive load, but in the case of a building air conditioner that uses the refrigeration air conditioner targeted by this embodiment, a heating load is generated at high outside temperatures. Since there is little to do, operation is performed at the initial set value as described above.
次に、主膨張弁8aの開度であるが、蒸発器となる空気熱交換器5aの出口であり、圧縮機3a吸入の状態(図2点E)の冷媒過熱度SHが、予め設定された目標値、例えば1℃となるように制御される。ここで空気熱交換器5aの出口であり圧縮機3a吸入の冷媒過熱度SHは、(温度センサ15a検知温度(圧縮機3の吸入温度))−(圧力センサ14aから換算される冷媒飽和温度)で演算される値を用いる(S37)。
主膨張弁8aの開度が小さくなると、空気熱交換器5aを流れる冷媒流量は減少し、空気熱交換器5a出口の冷媒過熱度SHは大きくなり、逆に主膨張弁8aの開度を大きくすると空気熱交換器5aの冷媒過熱度SHは小さくなる。そこで、圧縮機3a吸入(空気熱交換器5a出口)の冷媒過熱度SHと目標値とを比較し(S38)、冷媒過熱度SHが目標値より大きい場合には、主膨張弁8aの開度を大きく制御し、冷媒過熱度SHが目標値より小さい場合には主膨張弁8aの開度を小さく制御する(S39)。
Next, the opening degree of the main expansion valve 8a, which is the outlet of the air heat exchanger 5a serving as an evaporator, and the refrigerant superheat degree SH in the state of suction of the compressor 3a (point E in FIG. 2) is set in advance. The target value is controlled to be 1 ° C., for example. Here, the refrigerant superheat degree SH at the outlet of the air heat exchanger 5a and sucked into the compressor 3a is (temperature sensor 15a detected temperature (intake temperature of the compressor 3)) − (refrigerant saturation temperature converted from the pressure sensor 14a). (S37) is used.
When the opening of the main expansion valve 8a decreases, the flow rate of the refrigerant flowing through the air heat exchanger 5a decreases, the refrigerant superheat degree SH at the outlet of the air heat exchanger 5a increases, and conversely, the opening of the main expansion valve 8a increases. Then, the refrigerant superheat degree SH of the air heat exchanger 5a becomes small. Therefore, the refrigerant superheat degree SH of the compressor 3a intake (air heat exchanger 5a outlet) is compared with the target value (S38), and when the refrigerant superheat degree SH is larger than the target value, the opening degree of the main expansion valve 8a When the refrigerant superheat degree SH is smaller than the target value, the opening degree of the main expansion valve 8a is controlled to be small (S39).
次に、バイパス膨張弁10aの開度であるが、冷却運転と同様に行い、過冷却熱交換器7a出口の冷媒過熱度SHecoを演算し(S40)、この冷媒過熱度SHecoと目標値とを比較し(S41)、冷媒過熱度SHecoが目標値より大きい場合には、バイパス膨張弁10aの開度を大きく制御し、冷媒過熱度SHecoが目標値より小さい場合にはバイパス膨張弁10aの開度を小さく制御する(S42)。   Next, the opening of the bypass expansion valve 10a is performed in the same manner as the cooling operation, and the refrigerant superheat degree Sheco at the outlet of the supercooling heat exchanger 7a is calculated (S40), and the refrigerant superheat degree SHeco and the target value are calculated. In comparison (S41), when the refrigerant superheat degree SHeco is larger than the target value, the opening degree of the bypass expansion valve 10a is controlled to be large, and when the refrigerant superheat degree SHeco is smaller than the target value, the opening degree of the bypass expansion valve 10a. Is controlled to be small (S42).
なお、冷却・加熱運転におけるこれらの圧縮機3aの回転数制御や、主膨張弁8a、バイパス膨張弁10aの開度制御においては、目標値との偏差に基づくPID制御法などにより、制御量が決定される。   In the cooling / heating operation, the control amount is controlled by the PID control method based on the deviation from the target value in the rotational speed control of the compressor 3a and the opening control of the main expansion valve 8a and the bypass expansion valve 10a. It is determined.
次に、負荷状況に応じて、負荷側装置の運転状況が変化し、温水を送水するポンプ12の流量が変更される(S43)が、ポンプ流量が増減した場合の制御動作について説明する。まず、冷凍サイクル2a、2bともに動作している状況でポンプ流量、すなわち温水流量が増加した場合の運転動作について説明する。この場合、温水流量が増加する一方で、冷凍サイクル2a、2bの凝縮器熱交換量、即ち加熱能力は変化しないので、水熱交換器9出入口の温水温度差が減少し、水熱交換器9出口の温水温度が低下する。そこで温水出口温度が目標値となるように、各冷凍サイクル2の圧縮機3の回転数が増速される運転が行われる。   Next, the operation state of the load-side device changes according to the load state, and the flow rate of the pump 12 that supplies hot water is changed (S43), but the control operation when the pump flow rate increases or decreases will be described. First, an operation operation when the pump flow rate, that is, the hot water flow rate is increased while both the refrigeration cycles 2a and 2b are operating will be described. In this case, while the hot water flow rate is increased, the condenser heat exchange amount of the refrigeration cycles 2a and 2b, that is, the heating capacity does not change, so the temperature difference of the hot water at the inlet / outlet of the water heat exchanger 9 decreases, and the water heat exchanger 9 The hot water temperature at the outlet decreases. Therefore, an operation is performed in which the rotational speed of the compressor 3 of each refrigeration cycle 2 is increased so that the hot water outlet temperature becomes the target value.
逆に温水流量が減少した場合の運転動作は以下のようになる。この場合、温水流量が増加する一方で、冷凍サイクル2a、2bの凝縮器熱交換量、即ち加熱能力は変化しないので、水熱交換器9出入口の温水温度差が増加し、水熱交換器9出口の温水温度が上昇する。そこで温水出口温度が目標値となるように、各冷凍サイクル2の圧縮機3の回転数が減速される運転が行われる。   Conversely, the operation when the hot water flow rate decreases is as follows. In this case, while the hot water flow rate increases, the condenser heat exchange amount of the refrigeration cycles 2a and 2b, i.e., the heating capacity, does not change, so the hot water temperature difference at the inlet and outlet of the water heat exchanger 9 increases, and the water heat exchanger 9 The hot water temperature at the outlet rises. Therefore, an operation is performed in which the rotational speed of the compressor 3 of each refrigeration cycle 2 is decelerated so that the hot water outlet temperature becomes the target value.
圧縮機3の回転数については、動作下限値があるので、温水流量が、ある程度以上低下し、圧縮機3回転数が動作下限値に近づき、所定の回転数以下となったら、一方の冷凍サイクル2の圧縮機を停止し、運転させる冷凍サイクル2の数を減少させる。例えば、圧縮機3の下限回転数が20rpsの場合には、圧縮機3の回転数が30rpsとなった段階で、圧縮機の停止動作を行わせる。このとき、温水上流流路に接続される冷凍サイクルから順に停止動作を行う。冷凍サイクル2の停止に伴って、水熱交換器9出口の温水温度目標値が再設定される。この実施の形態では、冷凍サイクルの数が2から1に減少され、温水流路の最も下流側に位置する冷凍サイクル2aのみの運転となるので、この場合の温水出口温度目標値は、冷凍サイクルの数によらず45℃のままとなる。そして上述した冷凍サイクルの運転制御が実施される。   Since there is an operation lower limit value for the rotation speed of the compressor 3, when the flow rate of the hot water decreases more than a certain level and the rotation speed of the compressor 3 approaches the operation lower limit value and becomes a predetermined rotation speed or less, one refrigeration cycle 2 compressors are stopped and the number of refrigeration cycles 2 to be operated is reduced. For example, when the lower limit rotation speed of the compressor 3 is 20 rps, the compressor is stopped when the rotation speed of the compressor 3 reaches 30 rps. At this time, the stop operation is sequentially performed from the refrigeration cycle connected to the hot water upstream flow path. As the refrigeration cycle 2 stops, the hot water temperature target value at the outlet of the water heat exchanger 9 is reset. In this embodiment, since the number of refrigeration cycles is reduced from 2 to 1, and only the refrigeration cycle 2a located on the most downstream side of the hot water flow path is operated, the hot water outlet temperature target value in this case is the refrigeration cycle. It remains at 45 ° C. regardless of the number of. Then, the above-described operation control of the refrigeration cycle is performed.
冷凍サイクル2aのみ動作している状況で温水流量が減少した場合には、温水出口温度が目標値となるように、冷凍サイクル2aの圧縮機3aの回転数が減速される運転が行われる。さらに温水流量が減少し、圧縮機3aの回転数が下限値となるまで制御された場合には、下限回転数での運転を継続する。さらに温水流量が減少し、温水出口温度が目標値より所定温度、例えば1℃以上上回り、46℃以上となった場合には、冷凍サイクル2aの圧縮機3aも停止動作を行い、熱源機1の運転を停止する。   When the flow rate of hot water decreases while only the refrigeration cycle 2a is operating, an operation is performed in which the rotational speed of the compressor 3a of the refrigeration cycle 2a is decelerated so that the hot water outlet temperature becomes the target value. When the hot water flow rate is further decreased and the rotation speed of the compressor 3a is controlled until it reaches the lower limit value, the operation at the lower limit rotation speed is continued. Further, when the hot water flow rate is decreased and the hot water outlet temperature is higher than a target value by a predetermined temperature, for example, 1 ° C. or more, and becomes 46 ° C. or more, the compressor 3a of the refrigeration cycle 2a also performs a stop operation. Stop operation.
熱源機1が停止している状況で、温度センサ15hで検知される熱源機1の温水出口温度が負荷側への供給目標温度より、所定温度、例えば1℃以上下回り、44℃以下となった場合には、熱源機1の運転を行う。そして、温水流路の下流に配置される冷凍サイクル2aの圧縮機3aを起動し、水熱交換器9a出口の温水温度が目標値となるように圧縮機3aの回転数制御を行う。   While the heat source unit 1 is stopped, the hot water outlet temperature of the heat source unit 1 detected by the temperature sensor 15 h is lower than the target supply temperature to the load side by a predetermined temperature, for example, 1 ° C. or more and 44 ° C. or less. In that case, the heat source machine 1 is operated. And the compressor 3a of the refrigerating cycle 2a arrange | positioned downstream of a warm water flow path is started, and the rotation speed control of the compressor 3a is performed so that the warm water temperature of the water heat exchanger 9a exit may become a target value.
冷凍サイクル2aのみ動作している状況で温水流量が増加した場合には、温水出口温度が目標値となるように、冷凍サイクル2aの圧縮機3aの回転数が増速される運転が行われる。さらに温水流量が増加し、圧縮機3aの回転数が所定値以上となるまで制御された場合には、温水流路の上流に配置される冷凍サイクル2bの圧縮機3bを起動する。冷凍サイクル2bの運転に伴って、水熱交換器9出口の温水温度目標値が再設定され、各冷凍サイクルで水熱交換器9出口の温水温度が目標値となるように圧縮機の回転数制御が実施される。   When the hot water flow rate increases while only the refrigeration cycle 2a is operating, an operation is performed in which the rotational speed of the compressor 3a of the refrigeration cycle 2a is increased so that the hot water outlet temperature becomes the target value. When the flow rate of the hot water is further increased and the rotation speed of the compressor 3a is controlled to a predetermined value or more, the compressor 3b of the refrigeration cycle 2b disposed upstream of the hot water flow path is started. With the operation of the refrigeration cycle 2b, the hot water temperature target value at the outlet of the water heat exchanger 9 is reset, and the rotational speed of the compressor is set so that the hot water temperature at the outlet of the water heat exchanger 9 becomes the target value in each refrigeration cycle. Control is implemented.
次に、水熱交換器9の冷温水出口温度の目標設定と熱源機1の運転効率に関する特性を図5に基づいて説明する。図5は、冷凍サイクル2a、冷凍サイクル2bが同じ仕様、構成であり、各冷凍サイクルの圧縮機容量が20HPである場合に、熱源機1にて冷水を冷却する運転を行ったときの特性を表した図である。図の横軸は熱源機1全体の冷却熱交換量に対する冷温水下流側の配置される冷凍サイクル2aの冷却熱交換量の比率である。冷凍サイクル2a、2bの熱交換量が等しい場合は、熱交換量比率は0.5となる。冷水の熱交換においては、比熱が温度によらずほぼ一定となるので、熱交換量は水熱交換器9出入口の冷水温度差に比例する。従って図5横軸の熱交換量比率は、熱交換量比率≒水熱交換器9bの出入口の冷水温度差/熱源機1出入口の冷水温度となる。水熱交換器9bの出入口の冷水温度差は、運転制御により、水熱交換器9b出口の冷水温度目標値によって規定されることになり、水熱交換器9b出口の冷水温度目標値が高く設定されると、冷凍サイクル2bの熱交換量は少なくなり、冷凍サイクル2aの冷却熱交換量の比率は高くなる。逆に、水熱交換器9b出口の冷水温度目標値が低く設定されると、冷凍サイクル2bの熱交換量は多くなり、冷凍サイクル2aの冷却熱交換量の比率は低くなる。   Next, the characteristics regarding the target setting of the cold / hot water outlet temperature of the water heat exchanger 9 and the operation efficiency of the heat source unit 1 will be described with reference to FIG. FIG. 5 shows the characteristics when the refrigeration cycle 2a and the refrigeration cycle 2b have the same specifications and configuration, and when the compressor capacity of each refrigeration cycle is 20 HP, the operation of cooling the cold water in the heat source unit 1 is performed. FIG. The horizontal axis of the figure is the ratio of the cooling heat exchange amount of the refrigeration cycle 2a arranged on the downstream side of the cold / hot water to the cooling heat exchange amount of the entire heat source unit 1. When the heat exchange amounts of the refrigeration cycles 2a and 2b are equal, the heat exchange amount ratio is 0.5. In the heat exchange of the cold water, the specific heat is almost constant regardless of the temperature, so the heat exchange amount is proportional to the cold water temperature difference at the inlet / outlet of the water heat exchanger 9. Accordingly, the heat exchange rate ratio on the horizontal axis in FIG. 5 is: heat exchange rate ratio≈cold water temperature difference at the entrance / exit of the water heat exchanger 9b / cold water temperature at the entrance / exit of the heat source unit 1. The chilled water temperature difference at the inlet / outlet of the water heat exchanger 9b is defined by the chilled water temperature target value at the outlet of the water heat exchanger 9b by operation control, and the chilled water temperature target value at the outlet of the water heat exchanger 9b is set high. If it does, the heat exchange amount of the refrigerating cycle 2b will decrease, and the ratio of the cooling heat exchange amount of the refrigerating cycle 2a will become high. Conversely, when the target value of the cold water temperature at the outlet of the water heat exchanger 9b is set low, the amount of heat exchange in the refrigeration cycle 2b increases, and the ratio of the amount of cooling heat exchange in the refrigeration cycle 2a decreases.
図5には、熱源機1のCOP(運転効率)比が示されており、COP最大の状態を100%として比率を示している。図5には他に冷凍サイクル2a、2bの圧縮機3の回転数を表している。図5に示されているとおり、熱源機1のCOP比はサイクル2aの熱交換量比率=0.5付近をピークとした曲線となる。サイクル2aの熱交換量比率が0.4〜0.6の範囲では、COP比は最大効率の−2%程度以内に収まっており、この範囲で運転するようにすれば熱源機1を高効率で運転できる。   FIG. 5 shows the COP (operating efficiency) ratio of the heat source device 1, and the ratio is shown with the maximum COP state being 100%. FIG. 5 also shows the rotation speed of the compressor 3 of the refrigeration cycles 2a and 2b. As shown in FIG. 5, the COP ratio of the heat source device 1 is a curve having a peak at the heat exchange rate ratio of cycle 2a = 0.5. When the heat exchange amount ratio of the cycle 2a is in the range of 0.4 to 0.6, the COP ratio is within about -2% of the maximum efficiency. If the operation is performed in this range, the heat source unit 1 is highly efficient. You can drive in.
上記範囲で運転すると高効率運転できる理由は以下のように説明できる。冷凍サイクルの効率(COP)は一般に高温熱源と低温熱源の温度差が大きくなる程低下し、冷媒の温度で換算すると凝縮温度が高ければ高いほど低下し、蒸発温度が低ければ低いほど低下する。冷凍サイクル2aの冷却熱交換量の比率が0.5から大きく外れる場合、例えば冷凍サイクル2aの冷却熱交換量の比率が0.33の場合、冷凍サイクル2bの冷却熱交換量の比率は0.67となり、冷凍サイクル2bは冷凍サイクル2aの2倍の熱交換量で運転されることになる。
このとき、冷凍サイクル2aにおいては、冷却熱交換量=蒸発器の熱交換量が少ないため、蒸発器で冷水と冷媒が熱交換するための温度差が少なくなり、冷媒の蒸発温度は冷水温度と近接する。一方冷凍サイクル2bにおいては、冷却熱交換量=蒸発器の熱交換量が多く、蒸発器で冷水と冷媒が熱交換するための温度差が多く必要となり、冷媒の蒸発温度は冷水温度から離れて低下する。
The reason why high-efficiency operation is possible when operating in the above range can be explained as follows. The efficiency (COP) of the refrigeration cycle generally decreases as the temperature difference between the high temperature heat source and the low temperature heat source increases. When converted to the refrigerant temperature, the efficiency decreases (COP) decreases as the condensation temperature increases, and decreases as the evaporation temperature decreases. When the ratio of the cooling heat exchange amount of the refrigeration cycle 2a is significantly different from 0.5, for example, when the ratio of the cooling heat exchange amount of the refrigeration cycle 2a is 0.33, the ratio of the cooling heat exchange amount of the refrigeration cycle 2b is 0.00. Thus, the refrigeration cycle 2b is operated with a heat exchange amount twice that of the refrigeration cycle 2a.
At this time, in the refrigeration cycle 2a, the amount of cooling heat exchange = the amount of heat exchange of the evaporator is small, so the temperature difference for heat exchange between the cold water and the refrigerant in the evaporator is small, and the evaporation temperature of the refrigerant is equal to the cold water temperature. Proximity. On the other hand, in the refrigeration cycle 2b, the amount of cooling heat exchange = the amount of heat exchange in the evaporator is large, and a large temperature difference is required for heat exchange between the cold water and the refrigerant in the evaporator, and the evaporation temperature of the refrigerant is far from the cold water temperature. descend.
冷凍サイクル2aの冷却熱交換量の比率が0.5の場合の、冷水と冷媒蒸発温度との温度差をΔTとすると、冷凍サイクル2bの蒸発器では同一熱交換器仕様にて同一熱交換量で動作するので、冷凍サイクル2bの冷水と冷媒蒸発温度との温度差もΔTとなる。ここで前述したように、冷凍サイクル2aの冷却熱交換量の比率が0.33の場合を想定すると、冷凍サイクル2bの冷却熱交換量の比率は0.67となり、比率が0.5の場合に比べ熱交換量が33%増加し、冷凍サイクル2bは冷凍サイクル2aの2倍の熱交換量で運転されることになる。   When the temperature difference between the chilled water and the refrigerant evaporation temperature is ΔT when the ratio of the cooling heat exchange amount of the refrigeration cycle 2a is 0.5, the same heat exchanger specification is used for the evaporator of the refrigeration cycle 2b with the same heat exchanger specification. Therefore, the temperature difference between the cold water of the refrigeration cycle 2b and the refrigerant evaporation temperature is also ΔT. As described above, assuming that the ratio of the cooling heat exchange amount of the refrigeration cycle 2a is 0.33, the ratio of the cooling heat exchange amount of the refrigeration cycle 2b is 0.67, and the ratio is 0.5. As a result, the heat exchange amount is increased by 33%, and the refrigeration cycle 2b is operated at a heat exchange amount twice that of the refrigeration cycle 2a.
各水熱交換器9での冷水と冷媒蒸発温度との温度差は、冷凍サイクル2aでは蒸発器での熱交換量が熱交換量の比率が0.5の場合の2/3倍となるので、2/3×ΔT、冷凍サイクル2bでは蒸発器での熱交換量が熱交換量の比率が0.5の4/3倍となり、4/3×ΔTとなる。ここで、各冷凍サイクルのCOPを評価すると、熱交換量の比率が0.33の場合、冷凍サイクル2aのCOPは熱交換量の比率が0.5の場合に比べ、温度差1/3×ΔT分蒸発温度が上昇した分高くなり、冷凍サイクル2bのCOPは熱交換量の比率が0.5の場合に比べ、温度差1/3×ΔT分蒸発温度が低下した分低くなる。蒸発温度の変化が同じであるのでサイクル2aのCOP上昇幅とサイクル2bのCOP低下幅はほぼ等しくなる。   The temperature difference between the cold water and the refrigerant evaporation temperature in each water heat exchanger 9 is 2/3 times that in the refrigeration cycle 2a, the heat exchange amount in the evaporator is a ratio of the heat exchange amount of 0.5. In 2/3 × ΔT, in the refrigeration cycle 2b, the heat exchange rate in the evaporator is 4/3 times the ratio of the heat exchange rate, which is 4/3 × ΔT. Here, when the COP of each refrigeration cycle is evaluated, when the ratio of the heat exchange amount is 0.33, the COP of the refrigeration cycle 2a has a temperature difference of 1/3 × that of the case where the ratio of the heat exchange amount is 0.5. The COP of the refrigeration cycle 2b becomes higher as the evaporation temperature increases by ΔT, and the COP of the refrigeration cycle 2b becomes lower than the case where the ratio of heat exchange amount is 0.5, as the evaporation temperature decreases by 1/3 × ΔT. Since the change in the evaporation temperature is the same, the COP increase width of the cycle 2a and the COP decrease width of the cycle 2b are substantially equal.
冷凍サイクルの入力は入力=能力(蒸発器熱交換量)/COPで表されるので、COPが低ければ入力は増加するが、能力が増加するとそれに比例して入力も増加する。熱交換量の比率が0.33の場合、冷凍サイクル2bの能力は冷凍サイクル2aの2倍あるので、その分入力も増加する。従って、熱交換量の比率が0.33の場合の入力は、冷凍サイクル2aについては、(熱交換量の比率が0.5の場合の入力)×熱交換量比2/3×(1−温度差1/3×ΔT分のCOP上昇による効果)となり、冷凍サイクル2bについては、(熱交換量の比率が0.5の場合の入力)×熱交換量比4/3×(1+温度差1/3×ΔT分のCOP低下による効果)となる。熱源機1全体の入力は、両者を合算して、(熱交換量の比率が0.5の場合の入力)×2×(1+2/3×ΔT分のCOP低下による効果)となり、熱交換量の比率が0.5の場合の入力に比べ、熱源機1の入力は増加し、運転効率が低下する。   Since the input of the refrigeration cycle is expressed by input = capacity (evaporator heat exchange amount) / COP, the input increases if the COP is low, but the input increases proportionally as the capacity increases. When the ratio of the heat exchange amount is 0.33, the capacity of the refrigeration cycle 2b is twice that of the refrigeration cycle 2a, so the input increases accordingly. Therefore, the input when the heat exchange rate ratio is 0.33 is (input when the heat exchange rate ratio is 0.5) × heat exchange rate ratio 2/3 × (1− For refrigeration cycle 2b (input when heat exchange rate ratio is 0.5) × heat exchange rate ratio 4/3 × (1 + temperature difference) 1/3 × ΔT minute COP reduction effect). The total input of the heat source unit 1 is the sum of both (input when the ratio of heat exchange amount is 0.5) × 2 × (effect due to COP drop by 1 + 2/3 × ΔT), and the heat exchange amount Compared with the input when the ratio is 0.5, the input of the heat source unit 1 is increased and the operation efficiency is lowered.
熱交換量の比率が0.5の場合の入力に比べ、熱交換量の比率が0.33の場合は、冷凍サイクル2bの方の水熱交換器9が偏って負荷を賄う運転を行うことになり、その分各冷凍サイクル2の運転がアンバランスになり、効率が低下する。本実施の形態のように、水熱交換器9出入口の冷水温度の温度差が同じ程度になるように運転を行うことで、各冷凍サイクルの水熱交換器9がその容量に応じて、同じように負荷を賄う運転となり、各冷凍サイクル2がバランスよく運転を行うことができ、高効率運転を実現できる。   Compared to the input when the heat exchange rate is 0.5, when the heat exchange rate is 0.33, the water heat exchanger 9 in the refrigeration cycle 2b is biased to perform the operation to cover the load. Accordingly, the operation of each refrigeration cycle 2 becomes unbalanced, and the efficiency decreases. As in this embodiment, by operating so that the temperature difference of the cold water temperature at the inlet / outlet of the water heat exchanger 9 is the same, the water heat exchanger 9 of each refrigeration cycle is the same according to its capacity. Thus, the operation is to cover the load, and each refrigeration cycle 2 can operate in a well-balanced manner, thereby realizing a highly efficient operation.
また、本実施の形態では、水熱交換器9を冷水流路に対して直列に流すようにしたので高効率を得ることができる。図6は、水熱交換器9が1つしか無い場合(1蒸発回路)と、本実施の形態のように水熱交換器9が直列に配置されている場合(2蒸発回路)での、熱交換器での冷水と冷媒との温度差の特性を表した図である。図6では熱源機1の冷水入口温度が12℃、出口温度が7℃である場合の特性を表している。この場合の2蒸発回路での温度変化は図7のように表され、冷凍サイクル2bは蒸発温度ETbで運転され、冷凍サイクル2aは蒸発温度ETaで運転され、それぞれのサイクルの冷水と蒸発温度との温度差はΔTb、ΔTaとなる。上記実施の形態のように、各水熱交換器9での冷水出入口温度との温度差が同じく制御される場合にはΔTb=ΔTaとなる。このとき水熱交換器9aの冷水出入口温度は水熱交換器9bに比べ2.5℃低くなるので、冷凍サイクル2aの蒸発温度ETaは冷凍サイクル2bの蒸発温度ETbより2.5℃低くなる。   Moreover, in this Embodiment, since the water heat exchanger 9 was flowed in series with respect to the cold water flow path, high efficiency can be obtained. FIG. 6 shows a case where there is only one water heat exchanger 9 (1 evaporation circuit) and a case where the water heat exchanger 9 is arranged in series as in the present embodiment (2 evaporation circuit). It is a figure showing the characteristic of the temperature difference of the cold water and a refrigerant | coolant in a heat exchanger. FIG. 6 shows characteristics when the cold water inlet temperature of the heat source device 1 is 12 ° C. and the outlet temperature is 7 ° C. The temperature change in the two-evaporation circuit in this case is expressed as shown in FIG. 7, the refrigeration cycle 2b is operated at the evaporation temperature ETb, the refrigeration cycle 2a is operated at the evaporation temperature ETa, and the cold water and the evaporation temperature of each cycle are The difference in temperature is ΔTb and ΔTa. When the temperature difference with the cold water inlet / outlet temperature in each water heat exchanger 9 is similarly controlled as in the above embodiment, ΔTb = ΔTa. At this time, since the cold water inlet / outlet temperature of the water heat exchanger 9a is 2.5 ° C. lower than that of the water heat exchanger 9b, the evaporation temperature ETa of the refrigeration cycle 2a is 2.5 ° C. lower than the evaporation temperature ETb of the refrigeration cycle 2b.
図6の横軸は冷媒蒸発温度を表し、2蒸発回路では各冷凍サイクル2の平均値、(ETa+ETb)/2を表す。図6のグラフ(a)の縦軸温度差は、冷水と冷媒蒸発温度との対数平均温度差を表し、1蒸発回路の場合は、温度差=(冷水入口温度12℃−冷水出口温度7℃)/ln{(冷水入口温度12℃−冷媒蒸発温度)/(冷水出口温度7℃−冷媒蒸発温度)}となる。2蒸発回路の場合は、冷凍サイクル2aの温度差=(冷水入口温度9.5℃−冷水出口温度7℃)/ln{(冷水入口温度9.5℃−冷媒蒸発温度ETa)/(冷水出口温度7℃−冷媒蒸発温度ETa)}と、冷凍サイクル2bの温度差=(冷水入口温度12℃−冷水出口温度9.5℃)/ln{(冷水入口温度12℃−冷媒蒸発温度ETb)/(冷水出口温度9.5℃−冷媒蒸発温度ETb)}との平均値を表す。
図6のグラフ(b)の縦軸温度差は、上記のようにして求めて2蒸発回路の温度差−冷水と冷媒蒸発温度との対数平均温度差を表す。
The horizontal axis in FIG. 6 represents the refrigerant evaporation temperature, and in the 2-evaporation circuit, the average value of each refrigeration cycle 2 (ETa + ETb) / 2. The vertical axis temperature difference in the graph (a) of FIG. 6 represents the logarithm average temperature difference between the cold water and the refrigerant evaporation temperature. In the case of one evaporation circuit, the temperature difference = (cold water inlet temperature 12 ° C.−cold water outlet temperature 7 ° C. ) / Ln {(cold water inlet temperature 12 ° C.−refrigerant evaporation temperature) / (cold water outlet temperature 7 ° C.−refrigerant evaporation temperature)}. In the case of the two evaporation circuit, the temperature difference of the refrigeration cycle 2a = (cold water inlet temperature 9.5 ° C.−cold water outlet temperature 7 ° C.) / Ln {(cold water inlet temperature 9.5 ° C.−refrigerant evaporation temperature ETa) / (cold water outlet Temperature difference between temperature 7 ° C.−refrigerant evaporation temperature ETa)} and refrigeration cycle 2b = (cold water inlet temperature 12 ° C.—cold water outlet temperature 9.5 ° C.) / Ln {(cold water inlet temperature 12 ° C.−refrigerant evaporation temperature ETb) / (Cold water outlet temperature 9.5 ° C.−refrigerant evaporation temperature ETb)}.
The vertical axis temperature difference in the graph (b) of FIG. 6 is obtained as described above, and represents the temperature difference of the two evaporation circuits—the logarithm average temperature difference between the cold water and the refrigerant evaporation temperature.
図6に示されるように、冷水と冷媒蒸発温度との温度差は1蒸発回路に比べ、2蒸発回路の方が大きく、その分2蒸発回路の熱交換量は大きくなる。同一熱交換量の運転を行う場合は、1蒸発回路よりも2蒸発回路の方が蒸発温度を高く運転でき、その分高効率運転を行うことができる。また、2蒸発回路の温度差増加幅は、冷媒蒸発温度が高くなるほど拡大する。従って冷媒蒸発温度が高く、冷水と冷媒蒸発温度との温度差が小さくなる運転条件、即ち冷却負荷の小さい運転条件では、2蒸発回路での運転効率がより高くなる。   As shown in FIG. 6, the temperature difference between the cold water and the refrigerant evaporation temperature is larger in the 2-evaporation circuit than in the 1-evaporation circuit, and the heat exchange amount of the 2-evaporation circuit is increased accordingly. When the operation with the same heat exchange amount is performed, the two evaporation circuit can be operated at a higher evaporation temperature than the one evaporation circuit, and the highly efficient operation can be performed accordingly. In addition, the temperature difference increase width of the two evaporation circuit increases as the refrigerant evaporation temperature increases. Therefore, in the operating condition where the refrigerant evaporation temperature is high and the temperature difference between the cold water and the refrigerant evaporation temperature is small, that is, the operating condition where the cooling load is small, the operation efficiency in the two-evaporation circuit is higher.
図6に示される特性は以下の要因で生じる。一般に負荷側熱媒体は水など温度変化により顕熱で熱交換が行われる媒体が用いられる。このとき、冷水と冷媒との温度差が大きい場合、例えば水熱交換器9の冷水入口部では、熱交換量が大きくその分冷水の温度変化が大きくなる。一方、冷水と冷媒との温度差が小さい場合、例えば水熱交換器9の冷水出口部では、熱交換量が小さくその分冷水の温度変化が小さくなる。従って、水熱交換器9の内部では、冷水温度が高い領域よりも低い領域の方が一般に長くなる。この傾向は、高温部と低温部の温度差が大きいほど極端になり、冷水出口温度が冷媒蒸発温度と近接する運転条件となるほど、冷水温度が低い領域が長くなる。冷水温度が低い領域が長くなると、その分水熱交換器の伝熱性能に及ぼす影響が大きくなり、水熱交換器9の伝熱性能には、水熱交換器9の冷水出口温度と冷媒蒸発温度との温度差の影響を受けるようになる。   The characteristics shown in FIG. 6 are caused by the following factors. Generally, a medium such as water that exchanges heat with sensible heat due to a temperature change is used as the load-side heat medium. At this time, when the temperature difference between the cold water and the refrigerant is large, for example, at the cold water inlet portion of the water heat exchanger 9, the heat exchange amount is large and the temperature change of the cold water is correspondingly large. On the other hand, when the temperature difference between the cold water and the refrigerant is small, for example, at the cold water outlet of the water heat exchanger 9, the heat exchange amount is small and the temperature change of the cold water is correspondingly small. Accordingly, in the water heat exchanger 9, the region where the cold water temperature is low is generally longer than the region where the cold water temperature is high. This tendency becomes more extreme as the temperature difference between the high temperature part and the low temperature part is larger, and the region where the cold water temperature is lower becomes longer as the cold water outlet temperature becomes closer to the refrigerant evaporation temperature. When the region where the cold water temperature is low becomes longer, the influence on the heat transfer performance of the water heat exchanger increases accordingly, and the heat transfer performance of the water heat exchanger 9 includes the cold water outlet temperature and the refrigerant evaporation of the water heat exchanger 9. It will be affected by the temperature difference from the temperature.
ここで、冷媒蒸発温度が6℃の場合を評価すると、1蒸発回路の場合は、冷水出口温度を7℃とすると温度差が水熱交換器9の出口温度と冷媒蒸発温度との温度差が1℃であり、この温度差に応じた熱交換がなされる。一方2蒸発回路の場合は、平均蒸発温度6℃で運転されるとすると、各冷凍サイクル2の蒸発温度差が2.5℃存在するので、冷凍サイクル2aの蒸発温度ETaは4.75℃、冷凍サイクル2bの蒸発温度は7.25℃となる。このとき、冷水出口温度との温度差は、冷凍サイクル2a、2bとも2.25℃となり、1蒸発回路よりも大きくなる。この分、水熱交換器9での冷水と冷媒との温度差が拡大する。   Here, when evaluating the case where the refrigerant evaporation temperature is 6 ° C., in the case of one evaporation circuit, if the cold water outlet temperature is 7 ° C., the temperature difference is the difference between the outlet temperature of the water heat exchanger 9 and the refrigerant evaporation temperature. 1 ° C., and heat exchange is performed according to this temperature difference. On the other hand, in the case of the two-evaporation circuit, if it is operated at an average evaporation temperature of 6 ° C., the evaporation temperature difference of each refrigeration cycle 2 is 2.5 ° C., so the evaporation temperature ETa of the refrigeration cycle 2a is 4.75 ° C. The evaporation temperature of the refrigeration cycle 2b is 7.25 ° C. At this time, the temperature difference from the cold water outlet temperature is 2.25 ° C. for both the refrigeration cycles 2a and 2b, which is larger than that of the single evaporation circuit. Accordingly, the temperature difference between the cold water and the refrigerant in the water heat exchanger 9 increases.
この2蒸発回路による温度差拡大効果は、冷媒蒸発温度と冷水出口温度が近接しやすい条件、すなわち低負荷の条件であるほど効果が大きくなる。1蒸発回路では、低負荷となると、冷水出口温度と冷媒蒸発温度が近接するので、温度差が減少しやすくなり、低負荷による水熱交換器の熱交換量が低減した場合でも蒸発温度の上昇幅が小さくなり、効率向上効果が得られにくい。一方、2蒸発回路の場合は、熱交換量が減少した場合に、冷水出口温度と冷媒蒸発温度が近接しないので、低負荷による水熱交換器の熱交換量が低減した場合に、それに応じて蒸発温度を高く運転でき、効率向上効果が得られる。なお、この運転を実現するには、インバータによる圧縮機の容量制御が必須であり、逆に、インバータ圧縮機による容量制御と2蒸発回路を組み合わせることによって、低負荷時の大幅な効率向上効果が得られることになる。   The effect of expanding the temperature difference by the two-evaporation circuit increases as the refrigerant evaporation temperature and the chilled water outlet temperature are close to each other, that is, under a low load condition. In one evaporation circuit, when the load is low, the chilled water outlet temperature and the refrigerant evaporation temperature are close to each other, so the temperature difference is likely to decrease, and the evaporation temperature rises even when the heat exchange amount of the water heat exchanger due to the low load is reduced. The width becomes small, and it is difficult to obtain the efficiency improvement effect. On the other hand, in the case of the two-evaporation circuit, when the heat exchange amount decreases, the chilled water outlet temperature and the refrigerant evaporation temperature are not close to each other, so when the heat exchange amount of the water heat exchanger due to low load is reduced, The evaporating temperature can be increased and an efficiency improvement effect can be obtained. In order to realize this operation, capacity control of the compressor by the inverter is indispensable, and conversely, by combining capacity control by the inverter compressor and the two evaporation circuit, a significant efficiency improvement effect at low load can be achieved. Will be obtained.
また冷媒蒸発温度と冷水出口温度が近接しやすい他の条件として、いわゆる冷水の大温度差条件でも本実施の形態の効果は向上する。冷水が大温度差である場合、例えば冷水入口温度が17℃、出口温度が7℃で運転される場合を想定すると、前述の例にあげた冷水入口温度が12℃の場合に比べて、冷媒と冷水の平均的な温度差は大きくなるので水熱交換器9で同一熱交換量とした場合の冷媒蒸発温度はより高くなり、冷媒蒸発温度が冷水出口温度である7℃に近接するため、温度差向上に見合うほど蒸発温度は高く運転できない。従って1蒸発回路の場合には、大温度差条件で運転しても、大きな効率向上効果は得られない。   In addition, as another condition in which the refrigerant evaporation temperature and the cold water outlet temperature are likely to be close to each other, the effect of the present embodiment is improved even under a so-called cold water large temperature difference condition. When the cold water has a large temperature difference, for example, assuming that the cold water inlet temperature is 17 ° C. and the outlet temperature is 7 ° C., the refrigerant is smaller than the cold water inlet temperature shown in the above example is 12 ° C. And the average temperature difference between the chilled water and the water heat exchanger 9 makes the refrigerant evaporating temperature higher when the heat exchange amount is the same, and the refrigerant evaporating temperature is close to the chilled water outlet temperature of 7 ° C. The evaporation temperature is so high that the temperature difference can be improved. Therefore, in the case of a single evaporation circuit, a large efficiency improvement effect cannot be obtained even if the operation is performed under a large temperature difference condition.
一方、2蒸発回路の場合は、大温度差となると、上流側の水熱交換器9bでの冷水出口温度が上昇し、上記のように冷水入口温度が12℃から17℃になったとすると、水熱交換器9bでの冷水出口温度の目標値を9.5℃から12℃に変更して運転制御が実施される。この制御により、冷凍サイクル2bの蒸発温度ETbは水熱交換器9bでの冷水出口温度目標値の上昇に応じて高く運転されるので、高効率運転が可能となる。大温度差になればなるほど冷凍サイクル2bの蒸発温度ETbは高く運転されることになり、高効率運転が可能となる。   On the other hand, in the case of a two-evaporation circuit, when the temperature difference is large, the cold water outlet temperature in the upstream water heat exchanger 9b is increased, and the cold water inlet temperature is changed from 12 ° C to 17 ° C as described above. Operation control is performed by changing the target value of the cold water outlet temperature in the water heat exchanger 9b from 9.5 ° C to 12 ° C. By this control, the evaporating temperature ETb of the refrigeration cycle 2b is operated higher in accordance with the increase of the chilled water outlet temperature target value in the water heat exchanger 9b, so that highly efficient operation is possible. The larger the temperature difference is, the higher the evaporation temperature ETb of the refrigeration cycle 2b is operated, and the higher efficiency operation becomes possible.
なお、水熱交換器9の冷温水出口温度の目標設定については、負荷側熱媒体の物性に応じて補正を行ってもよい。負荷側熱媒体が水である場合は、温度による比熱の変化が小さいので、熱交換量は水熱交換器9での出入口温度差に比例するとできるが、ブラインなど他の媒体を用いる場合に、温度による比熱変化がある場合は、比熱変化に応じた補正を行ってもよい。   In addition, about the target setting of the cold / hot water outlet temperature of the water heat exchanger 9, you may correct | amend according to the physical property of a load side heat carrier. When the load-side heat medium is water, the change in specific heat due to temperature is small, so the heat exchange amount can be proportional to the inlet / outlet temperature difference in the water heat exchanger 9, but when using another medium such as brine, When there is a specific heat change due to temperature, correction according to the specific heat change may be performed.
水熱交換器9が複数ある場合の運転方法として、重要なポイントは、各水熱交換器9で熱負荷をその容量に応じてバランスよく賄うことにあるので、比熱の大きい温度条件では、それに応じて水熱交換器9での温度差を小さく、逆に比熱の小さい温度では水熱交換器9での温度差を大きく補正する。例えば、負荷側熱媒体として、エチレングリコールを用いた場合には、低温であるほど比熱が小さくなる。従って冷却運転時には、より低温であり水熱交換器9aでの温度差が水熱交換器9bでの温度差よりも、比熱が小さい分、それに反比例して大きくなるように、水熱交換器9出口の負荷側熱媒体の目標温度を設定する。   As an operation method when there are a plurality of water heat exchangers 9, an important point is that each water heat exchanger 9 covers the heat load in a well-balanced manner according to its capacity. Accordingly, the temperature difference in the water heat exchanger 9 is reduced, and conversely, the temperature difference in the water heat exchanger 9 is greatly corrected at a temperature with a small specific heat. For example, when ethylene glycol is used as the load-side heat medium, the specific heat decreases as the temperature decreases. Therefore, at the time of cooling operation, the water heat exchanger 9 has a lower temperature and the temperature difference in the water heat exchanger 9a is increased in inverse proportion to the smaller specific heat than the temperature difference in the water heat exchanger 9b. Set the target temperature of the load-side heat medium at the outlet.
なお、水熱交換器9の容量は、各冷凍サイクル2a、2bでアンバランスなく運転できるようにするため、基本的には圧縮機3の定格容量に応じた値に設計される。従って、冷温水出口温度の目標設定について、圧縮機3の定格容量を用いて、この容量に比例するように水熱交換器9での冷温水の温度差を設定し、冷温水出口温度の目標値を設定してもよい。この場合でも、水熱交換器9の容量に応じてバランスよく熱負荷を賄う運転を行うことができ、高効率の運転を実現できる。なお、一般に圧縮機3の定格容量は電動機出力が用いられるが、それ以外にも定格能力や、ストロークボリュームなど、定格容量に準じた値を適用してもよい。   The capacity of the water heat exchanger 9 is basically designed to a value corresponding to the rated capacity of the compressor 3 so that the refrigeration cycles 2a and 2b can be operated without imbalance. Therefore, for the target setting of the chilled / hot water outlet temperature, the temperature difference of the chilled / hot water in the water heat exchanger 9 is set in proportion to this capacity using the rated capacity of the compressor 3, and the target of the chilled / hot water outlet temperature is set. A value may be set. Even in this case, it is possible to perform an operation that covers the heat load in a well-balanced manner according to the capacity of the water heat exchanger 9, and a highly efficient operation can be realized. In general, the motor output is used as the rated capacity of the compressor 3, but other values such as rated capacity and stroke volume may be applied.
また圧縮機3の容量制御方法として以下のような運転を行ってもよい。まず、熱源機1の冷温水出口温度に相当する、温度センサ15gで検知される値が負荷側に供給する温度となるように、各冷凍サイクルの圧縮機3の合計容量を制御する。そしてその合計容量を各冷凍サイクルの圧縮機容量に分配するが、ここで、この容量分配は圧縮機3の定格容量に比例して分配する。圧縮機3a、3bが同容量の圧縮機の場合には同じ容量で分配され、インバータ圧縮機の場合は各圧縮機の運転回転数が同じ回転数になるように制御される。従って、冷却運転の場合は冷水出口温度が目標値より高いと各圧縮機の運転回転数を同じく上昇し、冷水出口温度が目標値より低いと各圧縮機の運転回転数を同じく低下させる運転を行い、加熱運転の場合は温水出口温度が目標値より低いと各圧縮機の運転回転数を同じく上昇し、温水出口温度が目標値より低いと各圧縮機の運転回転数を同じく低下させる運転を行う。   Further, as a capacity control method of the compressor 3, the following operation may be performed. First, the total capacity of the compressor 3 of each refrigeration cycle is controlled so that the value detected by the temperature sensor 15g corresponding to the cold / hot water outlet temperature of the heat source device 1 becomes the temperature supplied to the load side. The total capacity is distributed to the compressor capacity of each refrigeration cycle. This capacity distribution is distributed in proportion to the rated capacity of the compressor 3. When the compressors 3a and 3b are compressors having the same capacity, they are distributed with the same capacity, and when they are inverter compressors, the operation speed of each compressor is controlled to be the same. Therefore, in the case of cooling operation, if the chilled water outlet temperature is higher than the target value, the operating rotational speed of each compressor will be increased, and if the chilled water outlet temperature is lower than the target value, the operating speed of each compressor will be decreased. In the case of heating operation, if the hot water outlet temperature is lower than the target value, the operating rotational speed of each compressor will be increased, and if the hot water outlet temperature is lower than the target value, the operating speed of each compressor will be decreased. Do.
この運転の場合においても、圧縮機3の定格容量に応じて水熱交換器9の容量が設定されている場合には、各水熱交換器9において熱負荷を水熱交換器9の容量比に応じて賄うことになり、各冷凍サイクル2がバランスよく運転を行うことができ、高効率運転を実現できる。図5には、各冷凍サイクルの圧縮機回転数も示しているが、圧縮機回転数が等しくなる運転条件では、ほぼ最高効率に近い運転が実現されることが示されている。   Even in this operation, when the capacity of the water heat exchanger 9 is set according to the rated capacity of the compressor 3, the heat load is set to the capacity ratio of the water heat exchanger 9 in each of the water heat exchangers 9. Therefore, each refrigeration cycle 2 can be operated in a well-balanced manner, and high-efficiency operation can be realized. FIG. 5 also shows the compressor rotation speed of each refrigeration cycle, and it is shown that operation close to the maximum efficiency is realized under operating conditions where the compressor rotation speed is equal.
なお、図5に示されているように、各冷凍サイクルが同仕様かつ、圧縮機回転数が等しい運転条件であっても、各冷凍サイクルの水熱交換器9の熱交換量は等しくならず、水流路下流である冷凍サイクル2aの方の熱交換量が若干少ない運転となる。これは各冷凍サイクル2の運転蒸発温度の違いに起因し、圧縮機3は同一回転数で駆動されても、冷水温度が低いため運転蒸発温度が低くなる冷凍サイクル2aでは、その分圧縮機流量が低下し、水熱交換器9aでの熱交換量が低下する。従ってこの運転では、完全に各冷凍サイクルがバランスよく運転されているわけでなく、冷凍サイクル2bで賄われる熱負荷が若干多い運転となる。従って、この運転よりも冷凍サイクル2aで多く負荷が賄われるように、冷凍サイクル2aの圧縮機容量を高く設定するとより高効率に運転できる。   As shown in FIG. 5, the heat exchange amount of the water heat exchanger 9 in each refrigeration cycle is not equal even if each refrigeration cycle has the same specifications and operating conditions with the same compressor speed. The operation of the refrigeration cycle 2a downstream of the water flow path is slightly less. This is due to the difference in operating evaporating temperature of each refrigeration cycle 2, and even if the compressor 3 is driven at the same rotation speed, the refrigeration cycle 2a in which the operating evaporating temperature is low due to the low chilled water temperature reduces the compressor flow rate accordingly. Decreases, and the amount of heat exchange in the water heat exchanger 9a decreases. Therefore, in this operation, each refrigeration cycle is not completely operated in a well-balanced manner, and the heat load provided by the refrigeration cycle 2b is slightly increased. Therefore, if the compressor capacity of the refrigeration cycle 2a is set high so that the refrigeration cycle 2a can cover more load than this operation, the operation can be performed more efficiently.
また前述した制御で、各水熱交換器9での冷水温度差が、水熱交換器9での容量に応じた値となるように制御を行った場合には、水熱交換器9で賄う熱負荷はバランスよく運転される一方で、熱源側熱交換器である空気熱交換器5のバランスが若干崩れた運転となる。空気熱交換器5も、水熱交換器9と同様に一般に圧縮機3の定格容量に応じて設計されることになり、各冷凍サイクル2の水熱交換器9、および圧縮機3が同仕様である場合は、一般に空気熱交換器5も同仕様となる。この運転では、水熱交換器9の熱負荷はバランスよく運転されるが、冷凍サイクル2aは、冷凍サイクル2bに比べ低蒸発温度で運転されるので、その分運転効率COPが悪化し、圧縮機3の入力がより多くなる。空気熱交換器5での熱交換量は水熱交換器熱交換量+圧縮機入力となるので、この運転では、冷凍サイクル2aの空気熱交換器5の熱交換量が多い運転となり、その分、冷凍サイクル2aで賄われる熱負荷が多い運転となる。従って、この運転よりも冷凍サイクル2bで多く負荷が賄われるように、冷凍サイクル2bの水熱交換器9での冷水温度差を大きく設定し、冷凍サイクル2bの圧縮機容量が高くなるように運転するとより高効率に運転できる。   Further, in the above-described control, when the control is performed so that the chilled water temperature difference in each water heat exchanger 9 becomes a value corresponding to the capacity in the water heat exchanger 9, the water heat exchanger 9 covers it. While the heat load is operated in a well-balanced manner, the air heat exchanger 5 that is the heat source side heat exchanger is slightly out of balance. The air heat exchanger 5 is generally designed according to the rated capacity of the compressor 3 as with the water heat exchanger 9, and the water heat exchanger 9 and the compressor 3 of each refrigeration cycle 2 have the same specifications. In general, the air heat exchanger 5 has the same specifications. In this operation, the heat load of the water heat exchanger 9 is operated in a well-balanced manner, but the refrigeration cycle 2a is operated at a lower evaporation temperature than the refrigeration cycle 2b. 3 inputs are more. Since the heat exchange amount in the air heat exchanger 5 is the water heat exchanger heat exchange amount + the compressor input, in this operation, the air heat exchanger 5 in the refrigeration cycle 2a has a large heat exchange amount. This is an operation with a large heat load provided by the refrigeration cycle 2a. Therefore, the refrigeration cycle 2b is operated so that the compressor capacity of the refrigeration cycle 2b is increased by setting a large chilled water temperature difference in the water heat exchanger 9 of the refrigeration cycle 2b so that more load is covered by the refrigeration cycle 2b than in this operation. Then, it can drive more efficiently.
以上の特性から、冷凍サイクル2aの圧縮機容量が冷凍サイクル2bよりも高くなるように運転するとともに、冷凍サイクル2bの水熱交換器9bでの冷水温度差が冷凍サイクル2aよりも大きくなるように運転することで、前述した制御よりも、さらに高効率の運転を行うことができる。図5の特性においても、最大効率となる運転条件は、この条件にあてはまる領域にあることが示されている。   From the above characteristics, the compressor capacity of the refrigeration cycle 2a is operated so as to be higher than that of the refrigeration cycle 2b, and the chilled water temperature difference in the hydrothermal exchanger 9b of the refrigeration cycle 2b is larger than that of the refrigeration cycle 2a. By operating, it is possible to perform operation with higher efficiency than the control described above. Also in the characteristics of FIG. 5, it is shown that the operating condition that provides the maximum efficiency is in a region that satisfies this condition.
以上の水熱交換器9を冷水流路に直列に接続する効果は、冷却運転を例に説明したが、加熱運転時にも同様の効果が得られ、各冷凍サイクルの水熱交換器9が同じように加熱運転時の負荷を賄う運転を実施することで、各冷凍サイクル2がバランスよく運転を行うことができ、高効率運転を実現できる。また、低負荷運転時の効率向上効果も大きくなる。   The above-described effect of connecting the water heat exchanger 9 in series to the cold water flow path has been described by taking the cooling operation as an example, but the same effect can be obtained during the heating operation, and the water heat exchanger 9 of each refrigeration cycle is the same. Thus, by performing the operation which covers the load at the time of heating operation, each refrigeration cycle 2 can operate in a well-balanced manner, and a high-efficiency operation can be realized. Moreover, the efficiency improvement effect at the time of low load operation also becomes large.
また水熱交換器9を水流路に直列に接続する場合、下流側の水熱交換器9aについては、冷却運転の場合はより、低温の冷水が流入し、加熱運転の場合はより高温の温水が流入する。従って加熱運転時に冷媒の水熱交換器9出口温度を低く運転することができない運転となる。加熱運転の場合、水熱交換器9では冷媒側の熱は潜熱と併せて、冷媒のガス側、液側の顕熱を使って熱交換することになるが、下流側の冷凍サイクル2aについては、水熱交換器9が1つである場合に比べ、冷媒温度を十分に冷却できないことになり、冷媒の液側顕熱を活用できず、冷凍サイクルとしての効率が低下しやすい傾向となる。   Further, when the water heat exchanger 9 is connected in series to the water flow path, the downstream water heat exchanger 9a is supplied with low-temperature cold water in the cooling operation and hot water in the heating operation. Flows in. Accordingly, during the heating operation, the refrigerant water heat exchanger 9 outlet temperature cannot be lowered. In the case of heating operation, in the water heat exchanger 9, the heat on the refrigerant side is combined with latent heat and heat exchange is performed using sensible heat on the gas side and liquid side of the refrigerant. As compared with the case where there is only one water heat exchanger 9, the refrigerant temperature cannot be sufficiently cooled, the liquid side sensible heat of the refrigerant cannot be utilized, and the efficiency as the refrigeration cycle tends to be lowered.
そこで、本実施の形態では、そのデメリットを補うべく、エコノマイザ回路を搭載し、冷媒の液側顕熱を活用できる構成としている。即ち、過冷却熱交換器7で熱交換することで、高圧の液冷媒の顕熱を使って中間圧の冷媒を加熱し、この冷媒を圧縮し圧縮機3で圧縮し、水熱交換器9に流入させることで、間接的に冷媒の液側顕熱を水熱交換器9で温水を加熱するための熱として用いている。これにより、水熱交換器9を水流路に直列に接続する場合であっても、加熱運転時に冷媒の液側顕熱を十分に活用でき、より高効率の運転を行うことができる。   Therefore, in the present embodiment, an economizer circuit is mounted to compensate for the disadvantages, and the liquid side sensible heat of the refrigerant can be utilized. That is, heat exchange is performed by the supercooling heat exchanger 7 to heat the intermediate pressure refrigerant using the sensible heat of the high pressure liquid refrigerant, the refrigerant is compressed and compressed by the compressor 3, and the water heat exchanger 9 The liquid-side sensible heat of the refrigerant is indirectly used as heat for heating the hot water in the water heat exchanger 9. Thereby, even if it is a case where the water heat exchanger 9 is connected to a water flow path in series, the liquid side sensible heat of a refrigerant | coolant can fully be utilized at the time of a heating operation, and a more efficient driving | operation can be performed.
また、本実施の形態では、水熱交換器9の流路は、冷却運転時は冷媒と冷水が並行して流れる並行流となり、加熱運転時は冷媒と温水が対向して流れる対向流となるように流路構成される。一般に冷温水を供給するチラーでは、冷却運転時に水熱交換器9で冷媒が冷水と対向し流れる対向流の構成とされることが多い。通常1蒸発回路である従来例では、冷却運転の効率向上をねらったときに、大温度差条件や低負荷条件などで、冷媒温度と冷水出口温度が近接する運転において、対向流構成としておくと、前述したような蒸発温度上昇させる効果は望みにくいが、冷媒出口側の蒸発温度と冷水入口温度との温度差が大きいので、この温度差を活かして、水熱交換器9の冷媒出口部で過熱度SHを大きくとることができ、それによる効率向上が実現できていた。   In the present embodiment, the flow path of the water heat exchanger 9 is a parallel flow in which the refrigerant and the cold water flow in parallel during the cooling operation, and a counter flow in which the refrigerant and the hot water flow in the heating operation. The flow path is configured as follows. In general, a chiller that supplies cold / hot water is often configured to have a counterflow structure in which the refrigerant flows in the hydrothermal exchanger 9 to face the cold water during the cooling operation. In the conventional example, which normally has one evaporation circuit, when the efficiency of the cooling operation is aimed to be improved, a counter flow configuration is used in an operation in which the refrigerant temperature and the cold water outlet temperature are close to each other under a large temperature difference condition or a low load condition. Although the effect of raising the evaporation temperature as described above is difficult to expect, since the temperature difference between the evaporation temperature on the refrigerant outlet side and the cold water inlet temperature is large, this temperature difference is utilized to make the refrigerant outlet portion of the water heat exchanger 9 The degree of superheat SH can be increased, and the efficiency can be improved accordingly.
本実施の形態のように、2つの水熱交換器9a、9bを直列に配置した構成とする場合、冷却運転において、大温度差条件や低負荷条件となっても、冷水冷媒出口側の蒸発温度と冷水入口温度との温度差は小さいため、水熱交換器9の冷媒出口部で過熱度SHを大きくとる運転は実施できない。従って冷却運転では、冷媒温度がほぼ一定の蒸発温度である状況で水と熱交換することになるので、水熱交換器9の流路が対向流であっても並行流であってもほぼ同程度の熱交換性能となる。
加熱運転においても、水熱交換器9での冷媒出口側の凝縮温度と温水入口温度との温度差は小さく、冷媒出口部で過冷却度SCを大きくとれない運転となる。しかし、加熱運転では冷媒入口が過熱ガスであり、冷媒凝縮温度よりも高温である。この高温部の冷媒で、低温である水熱交換器9入口部の温水を加熱するよりは、高温である水熱交換器9出口部の温水を加熱する方が、熱ロスを少なく高温冷媒を活用でき、水熱交換器の性能を高くできる。そこで、本実施の形態のように2つの水熱交換器9a,9bを直列に配置した構成とする場合、加熱運転時は冷媒と温水が対向して流れる対向流となるように流路構成とすることで、高効率運転を実現できる。
When the two water heat exchangers 9a and 9b are arranged in series as in the present embodiment, even if a large temperature difference condition or a low load condition occurs in the cooling operation, the evaporation on the cold water refrigerant outlet side Since the temperature difference between the temperature and the cold water inlet temperature is small, an operation for increasing the degree of superheat SH at the refrigerant outlet of the water heat exchanger 9 cannot be performed. Therefore, in the cooling operation, heat exchange with water is performed in a situation where the refrigerant temperature is a substantially constant evaporation temperature. Therefore, the flow path of the water heat exchanger 9 is almost the same regardless of whether it is a counter flow or a parallel flow. The heat exchange performance is of the order.
Even in the heating operation, the temperature difference between the condensing temperature on the refrigerant outlet side in the water heat exchanger 9 and the hot water inlet temperature is small, and the supercooling degree SC cannot be increased at the refrigerant outlet portion. However, in the heating operation, the refrigerant inlet is superheated gas, which is higher than the refrigerant condensation temperature. Rather than heating the hot water at the inlet of the low-temperature water heat exchanger 9 with this high-temperature refrigerant, heating the hot water at the outlet of the high-temperature water heat exchanger 9 reduces heat loss and reduces the high-temperature refrigerant. It can be used and the performance of the water heat exchanger can be improved. Therefore, when the two water heat exchangers 9a and 9b are arranged in series as in the present embodiment, the flow path configuration is such that the refrigerant and the hot water flow in opposite directions during the heating operation. By doing so, high efficiency operation can be realized.
また、本実施の形態では、動作する冷凍サイクルの数が減少する場合には、流路下流側の冷凍サイクル2aが動作するように制御を行う。停止した冷凍サイクル2bでは、圧縮機3bの停止により冷媒が流れないため、水熱交換器9bでの熱交換量は減少するが、それでも水熱交換器9bを冷温水が流れることにより、いくらかの熱交換がなされる。このとき冷媒温度は、熱源機1周囲の温度とおおよそ等しくなるので、一般に冷却運転時に冷媒は加熱され、温水運転時には冷却され、若干の熱ロスが生じる。このとき、停止する冷凍サイクル2が冷温水流路下流にあると、運転している冷凍サイクル2の冷温水出口温度は、熱ロスに見合った分、冷却運転時は低下、加熱運転時は上昇させる運転となる。冷凍サイクル2の運転数が減少する場合、基本的に負荷が少ない運転となるので、冷温水出口温度と冷媒蒸発温度、冷媒凝縮温度が近接する運転となる。従って、熱ロスにより冷温水出口温度目標を変動しなければならない運転となると、その変動分、冷媒蒸発温度を低く、冷媒凝縮温度を高く運転しなければならず、運転効率が低下する。   In the present embodiment, when the number of operating refrigeration cycles decreases, control is performed so that the refrigeration cycle 2a on the downstream side of the flow path operates. In the stopped refrigeration cycle 2b, since the refrigerant does not flow due to the stop of the compressor 3b, the amount of heat exchange in the water heat exchanger 9b is reduced, but there is still some amount of cold / hot water flowing through the water heat exchanger 9b. Heat exchange is performed. At this time, since the refrigerant temperature is approximately equal to the temperature around the heat source unit 1, the refrigerant is generally heated during the cooling operation and cooled during the hot water operation, causing a slight heat loss. At this time, if the refrigeration cycle 2 to be stopped is downstream of the cold / hot water flow path, the cold / hot water outlet temperature of the refrigeration cycle 2 being operated is lowered during the cooling operation and raised during the heating operation according to the heat loss. It becomes driving. When the number of operations of the refrigeration cycle 2 is reduced, the operation is basically performed with a small load, and therefore, the operation is such that the cold / hot water outlet temperature is close to the refrigerant evaporation temperature and the refrigerant condensation temperature. Therefore, when the operation in which the cold / hot water outlet temperature target has to be changed due to heat loss is required, the refrigerant evaporating temperature must be lowered and the refrigerant condensing temperature must be increased, and the operating efficiency is lowered.
一方、本実施の形態のように、停止する冷凍サイクル2が冷温水流路上流にあると、熱ロス分だけ、運転している冷凍サイクル2に流入する冷温水入口温度が変動するが、冷温水出口温度は熱源機1から負荷側に供給する温度となる。従って、熱ロスが生じても、冷温水出口温度目標はそのまま運転でき、停止する冷凍サイクル2が冷温水流路下流にある場合の熱ロスによる運転効率低下が生じず、高効率の運転を実現できる。
なお、冷凍サイクル2の動作停止時の熱ロスを抑制するための対策として、水熱交換器9での冷温水の流路をバイパスする流路を設け、冷凍サイクル2停止時は、冷温水がバイパス流路を流れる構成としてもよいし、冷凍サイクル2停止時に水熱交換器9を冷媒が流れないように、水熱交換器9前後を逆止弁などの弁類で閉止する機構を設けてもよい。これらの対策により熱ロスを抑制でき、高効率の運転を実現できる。
On the other hand, if the refrigeration cycle 2 to be stopped is upstream of the cold / hot water flow path as in the present embodiment, the temperature of the cold / hot water inlet flowing into the refrigeration cycle 2 that is operating varies by the amount of heat loss. The outlet temperature is the temperature supplied from the heat source unit 1 to the load side. Therefore, even if a heat loss occurs, the cold / hot water outlet temperature target can be operated as it is, and when the refrigeration cycle 2 to be stopped is downstream of the cold / hot water flow path, the operation efficiency is not lowered due to the heat loss, and a highly efficient operation can be realized. .
As a measure for suppressing the heat loss when the operation of the refrigeration cycle 2 is stopped, a flow path that bypasses the flow path of the cold / hot water in the water heat exchanger 9 is provided. It is good also as a structure which flows through a bypass flow path, and provided the mechanism which closes the water heat exchanger 9 front and back with valves, such as a check valve, so that a refrigerant may not flow through the water heat exchanger 9 at the time of the refrigerating cycle 2 stop. Also good. These measures can suppress heat loss and realize highly efficient operation.
なお、熱源機1に搭載される冷凍サイクル2の数は2個に限定されるものではなく、3個以上搭載してもよい。この場合も各冷凍サイクルの水熱交換器9を負荷側熱媒体に対して直列となるように構成するとともに、冷温水の温度差が搭載される水熱交換器9の容量、もしくは圧縮機3の容量に比例するように運転することで高効率の運転を実現できる。   Note that the number of refrigeration cycles 2 mounted on the heat source device 1 is not limited to two, and three or more refrigeration cycles may be mounted. In this case as well, the water heat exchanger 9 of each refrigeration cycle is configured in series with the load-side heat medium, and the capacity of the water heat exchanger 9 in which the temperature difference of the cold / hot water is mounted, or the compressor 3 High-efficiency operation can be achieved by operating in proportion to the capacity.
なお、圧縮機3の容量制御方法については、インバータによる回転数制御だけでなく、他の手法を用いてもよい。例えば機械的に圧縮機3のストロークボリュームを変更する容量制御方法を用いてもよいし、圧縮機3を複数台設け、その運転台数を変更することで圧縮機3の容量制御を行ってもよい。この場合も、インバータで制御する場合と同様の運転制御を行うことで、各冷凍サイクル2の運転をバランスよく実施でき、高効率の運転を実現できる。   In addition, about the capacity | capacitance control method of the compressor 3, you may use not only the rotation speed control by an inverter but another method. For example, a capacity control method that mechanically changes the stroke volume of the compressor 3 may be used, or a plurality of compressors 3 may be provided, and the capacity control of the compressor 3 may be performed by changing the number of operating units. . Also in this case, by performing the same operation control as that controlled by the inverter, the operation of each refrigeration cycle 2 can be carried out in a balanced manner, and a highly efficient operation can be realized.
また水熱交換器9の構成については、プレート熱交換器に限定するものでは無く、他の構成、例えばシェルチューブ型や、二重管式などの構成を用いてもよい。空気熱交換器5についても、プレートフィン熱交換器に限定されるものでなく、コルゲートフィンを用いるなど他の形式を用いてもよい。また熱源側熱交換器としては空気を媒体とするものだけでなく、水など他の媒体を用いるものを適用しても同様の効果を実現できる。   Moreover, about the structure of the water heat exchanger 9, it is not limited to a plate heat exchanger, For example, you may use structures, such as a shell tube type and a double pipe type. The air heat exchanger 5 is not limited to the plate fin heat exchanger, and other types such as corrugated fins may be used. Further, the same effect can be realized by applying not only a heat source side heat exchanger using air as a medium but also one using other medium such as water.
また負荷側熱媒体としては、水やブラインだけでなく、顕熱で熱を授受する媒体であれば他の媒体でも同様の効果を得ることができる。例えば空気を用いる場合であっても、各負荷側熱交換器を風路構成内に直列に配置することで、高効率運転を実現できる。
冷媒としては、R410Aを例に説明したが、他の冷媒、例えばR407C、R404A、NH3、CO2などであっても同様の効果を得ることができる。本実施の形態では、潜熱で温度を伝えるとともに、その間温度一定である媒体を適用する場合の高効率化手段として有効である。R407Cなど潜熱で温度を伝える場合に温度変化する媒体を適用すると、冷媒温度と冷水出口温度が近接する状況となりにくく、効果は小さくなるが、潜熱で温度を伝える場合に温度変化しない媒体で、特に高圧冷媒であり圧力損失による温度変化が小さいR410AやCO2を適用した場合には、冷媒温度と冷水出口温度が近接する状況となりやすく、本実施の形態を適用した場合の効果が高まる。
As the load-side heat medium, not only water and brine, but also other media can obtain the same effect as long as the medium transfers heat by sensible heat. For example, even when air is used, high efficiency operation can be realized by arranging each load side heat exchanger in series in the air path configuration.
Although R410A has been described as an example of the refrigerant, other refrigerants such as R407C, R404A, NH3, CO2 and the like can achieve the same effect. In the present embodiment, the temperature is transmitted by latent heat and is effective as a high efficiency means when applying a medium in which the temperature is constant during that time. Applying a medium that changes temperature when transferring temperature by latent heat, such as R407C, makes it difficult for the refrigerant temperature and the chilled water outlet temperature to be close to each other, and the effect is reduced. When R410A or CO 2 that is a high-pressure refrigerant and has a small temperature change due to pressure loss is applied, the refrigerant temperature and the cold water outlet temperature are likely to be close to each other, and the effect of applying this embodiment is enhanced.
実施の形態2.
以下この発明の実施の形態2を図8に示す。図8は実施の形態2における熱源機1の冷媒回路構成を表したものであり、冷凍サイクル2bにおいて、圧縮機3bと並列に圧縮機3cが設けられている。圧縮機3cは、圧縮機3a、3bと同仕様のインバータ圧縮機である。圧縮機3cが設けてある冷凍サイクル2bは、冷凍サイクル2aの2倍の圧縮機容量となり、それに応じて空気熱交換器5b、水熱交換器9bの伝熱面積や、空気熱交換器5bのファン風量は冷凍サイクル2aの2倍に設定される。これ以外の構成については、実施の形態1と同様であるので説明を省略する。
Embodiment 2. FIG.
A second embodiment of the present invention is shown in FIG. FIG. 8 illustrates a refrigerant circuit configuration of the heat source apparatus 1 according to the second embodiment. In the refrigeration cycle 2b, a compressor 3c is provided in parallel with the compressor 3b. The compressor 3c is an inverter compressor having the same specifications as the compressors 3a and 3b. The refrigeration cycle 2b provided with the compressor 3c has a compressor capacity twice that of the refrigeration cycle 2a, and accordingly, the heat transfer area of the air heat exchanger 5b and the water heat exchanger 9b, the air heat exchanger 5b, The fan air volume is set to twice that of the refrigeration cycle 2a. Since other configurations are the same as those in the first embodiment, description thereof is omitted.
実施の形態2における装置の動作、および容量制御方法も実施の形態1と同様に実施される。水熱交換器9での冷温水出口温度の目標設定については、水熱交換器9bの容量が水熱交換器9aの2倍であること、または冷凍サイクル2bに搭載されている圧縮機3の容量が冷凍サイクル2aに搭載されている容量の2倍であることから、冷凍サイクル2bの冷温水温度差が冷凍サイクル2aの冷温水温度差の2倍となるように設定される。   The operation of the apparatus and the capacity control method in the second embodiment are also performed in the same manner as in the first embodiment. Regarding the target setting of the cold / hot water outlet temperature in the water heat exchanger 9, the capacity of the water heat exchanger 9b is twice that of the water heat exchanger 9a, or the compressor 3 mounted in the refrigeration cycle 2b. Since the capacity is twice the capacity installed in the refrigeration cycle 2a, the cold / hot water temperature difference of the refrigeration cycle 2b is set to be twice the cold / hot water temperature difference of the refrigeration cycle 2a.
例えば、冷却運転においては、熱源機1に流入する冷水温度が12℃、供給される冷水の目標温度が7℃の場合には、熱源機1全体の冷水温度差が12−7=5℃となるため、冷凍サイクル2aの冷水温度差は5×1/3=1.67℃、冷凍サイクル2bの冷水温度差が5×2/3=3.33℃となるように設定される。従って、水熱交換器9が冷水流路上流に配置される冷凍サイクル2bでは、水熱交換器9b出口の冷水温度目標値は12−3.33=8.67℃に設定される、水熱交換器9が冷水流路下流に配置される冷凍サイクル2aでは、水熱交換器9a出口の冷水温度目標値は8.67−1.67=7℃に設定され、これは当然負荷側装置に供給する冷水温度の目標値と等しくなる。   For example, in the cooling operation, when the temperature of the chilled water flowing into the heat source unit 1 is 12 ° C. and the target temperature of the supplied chilled water is 7 ° C., the chilled water temperature difference of the entire heat source unit 1 is 12−7 = 5 ° C. Therefore, the cold water temperature difference of the refrigeration cycle 2a is set to 5 × 1/3 = 1.67 ° C., and the cold water temperature difference of the refrigeration cycle 2b is set to 5 × 2/3 = 3.33 ° C. Therefore, in the refrigeration cycle 2b in which the water heat exchanger 9 is arranged upstream of the cold water flow path, the target value of the cold water temperature at the outlet of the water heat exchanger 9b is set to 12−3.33 = 8.67 ° C. In the refrigeration cycle 2a in which the exchanger 9 is arranged downstream of the chilled water flow path, the chilled water temperature target value at the outlet of the water heat exchanger 9a is set to 8.67-1.67 = 7 ° C. It becomes equal to the target value of the supplied cold water temperature.
加熱運転においては、例えば、熱源機1に流入する温水温度が40℃、供給される温水の目標温度が45℃の場合には、熱源機1全体の温水温度差が45−40=5℃となるため、冷凍サイクル2aの温水温度差は1.67℃、冷凍サイクル2bは3.33℃となるように設定され、水熱交換器9が温水流路上流に配置される冷凍サイクル2bでは、水熱交換器9b出口の温水温度目標値は40+3.33=43.33℃に設定される。水熱交換器9が温水流路下流に配置される冷凍サイクル2aでは、水熱交換器9a出口の温水温度目標値は43.33+1.67=45℃に設定され、これは当然負荷側装置に供給する温水温度の目標値と等しくなる。   In the heating operation, for example, when the temperature of hot water flowing into the heat source unit 1 is 40 ° C. and the target temperature of the supplied hot water is 45 ° C., the temperature difference of the hot water source 1 as a whole is 45−40 = 5 ° C. Therefore, in the refrigeration cycle 2b in which the hot water temperature difference of the refrigeration cycle 2a is set to 1.67 ° C., the refrigeration cycle 2b is set to 3.33 ° C., and the water heat exchanger 9 is arranged upstream of the hot water flow path, The target value of the hot water temperature at the outlet of the water heat exchanger 9b is set to 40 + 3.33 = 43.33 ° C. In the refrigeration cycle 2a in which the water heat exchanger 9 is arranged downstream of the hot water flow path, the target value of the hot water temperature at the outlet of the water heat exchanger 9a is set to 43.33 + 1.67 = 45 ° C. It becomes equal to the target value of the hot water temperature to be supplied.
このように冷温水の出口温度目標を設定することで、各冷凍サイクルの水熱交換器9が水熱交換器9の容量に応じて、同じように負荷を賄う運転となり、各冷凍サイクル2a、2bをバランスよく運転を行うことができ、高効率運転を実現できる。   By setting the outlet temperature target of the cold / hot water in this way, the water heat exchanger 9 of each refrigeration cycle is operated to cover the load in the same manner according to the capacity of the water heat exchanger 9, and each refrigeration cycle 2a, 2b can be operated with good balance, and high-efficiency operation can be realized.
また、各冷凍サイクルの冷温水出口温度目標を設定する代わりに、各圧縮機3の回転数が同一となるように運転制御を行ってもよい。この場合、圧縮機3の定格容量に応じて水熱交換器9の容量が設定されているので、各水熱交換器9において熱負荷を水熱交換器9の容量比に応じて賄うことになり、各冷凍サイクル2がバランスよく運転を行うことができ、高効率運転を実現できる。   Further, instead of setting the cold / hot water outlet temperature target of each refrigeration cycle, the operation control may be performed so that the rotation speeds of the compressors 3 are the same. In this case, since the capacity of the water heat exchanger 9 is set according to the rated capacity of the compressor 3, the heat load in each of the water heat exchangers 9 is covered according to the capacity ratio of the water heat exchanger 9. Thus, each refrigeration cycle 2 can be operated in a balanced manner, and high-efficiency operation can be realized.
なお、本実施の形態では、冷凍サイクル2に搭載される圧縮機3の容量が大きい冷凍サイクル2bを冷温水流路の上流側に配置している。冷凍サイクル2a、2bの運転を比較すると、冷却運転の場合、流入する冷水温度は冷凍サイクル2bの方が高く、加熱運転の場合、流入する温水温度は冷凍サイクル2bの方が低くなり、その分冷凍サイクル2bの運転効率が高くなる。熱源機1に搭載される圧縮機3の容量を大きくする場合、運転効率の低い冷凍サイクル2の容量を増加させるよりは、運転効率の高い冷凍サイクル2の容量を増加させる方が、熱源機1全体としての入力増加が少なく、高効率となる。従って、本実施の形態のように、各冷凍サイクル2に搭載される圧縮機3の容量が異なる場合には、容量の大きい圧縮機3が搭載される冷凍サイクル2を冷温水流路の上流側に配置することで、より高効率の装置とすることができる。   In the present embodiment, the refrigeration cycle 2b having a large capacity of the compressor 3 mounted in the refrigeration cycle 2 is arranged on the upstream side of the cold / hot water flow path. Comparing the operation of the refrigeration cycles 2a and 2b, in the cooling operation, the inflowing chilled water temperature is higher in the refrigeration cycle 2b, and in the heating operation, the inflowing hot water temperature is lower in the refrigeration cycle 2b. The operating efficiency of the refrigeration cycle 2b is increased. When the capacity of the compressor 3 mounted on the heat source apparatus 1 is increased, it is more likely to increase the capacity of the refrigeration cycle 2 with high operating efficiency than to increase the capacity of the refrigeration cycle 2 with low operating efficiency. There is little increase in input as a whole, and high efficiency is achieved. Therefore, when the capacity of the compressor 3 mounted in each refrigeration cycle 2 is different as in this embodiment, the refrigeration cycle 2 mounted with the large capacity compressor 3 is placed upstream of the cold / hot water flow path. By arranging, a more efficient device can be obtained.
実施の形態3.
以下この発明の実施の形態3を図9に示す。実施の形態3では実施の形態1におけるエコノマイザ回路の変わりに、過冷却熱交換器7にて圧縮機3の吸入冷媒と高圧の液冷媒を熱交換する構成としている。図3の他の構成、および運転制御については実施の形態1と同様である。
Embodiment 3 FIG.
A third embodiment of the present invention is shown in FIG. In the third embodiment, instead of the economizer circuit in the first embodiment, the supercooling heat exchanger 7 exchanges heat between the suction refrigerant of the compressor 3 and the high-pressure liquid refrigerant. Other configurations in FIG. 3 and operation control are the same as those in the first embodiment.
本実施の形態では、高圧液冷媒の顕熱を使って圧縮機3の吸入冷媒を加熱し、この冷媒を圧縮機3で圧縮し、水熱交換器9に流入させることで、間接的に冷媒の液側顕熱を水熱交換器9で温水を加熱するための熱として用いている。これにより、水熱交換器9を水流路に直列に接続する場合であっても、加熱運転時に冷媒の液側顕熱を十分に活用でき、より高効率の運転を行うことができる。   In the present embodiment, the sensible heat of the high-pressure liquid refrigerant is used to heat the refrigerant sucked in the compressor 3, the refrigerant is compressed by the compressor 3, and flows into the water heat exchanger 9. The liquid side sensible heat is used as heat for heating the hot water in the water heat exchanger 9. Thereby, even if it is a case where the water heat exchanger 9 is connected to a water flow path in series, the liquid side sensible heat of a refrigerant | coolant can fully be utilized at the time of a heating operation, and a more efficient driving | operation can be performed.
実施の形態4.
以下この発明の実施の形態4を説明する。実施の形態4では実施の形態1と同じ図1の回路構成とする。加熱運転時は空気条件によっては、空気熱交換器5に着霜が生じるので、デフロスト運転を実施する必要がある。一般にデフロスト運転を実施するときは加熱運転を停止し、負荷側への熱供給を停止して、デフロスト運転を実施し、終了後加熱運転を再開する。従って、デフロスト運転時は、熱負荷を賄えず、空調運転の快適性上問題となる場合があった。
Embodiment 4 FIG.
Embodiment 4 of the present invention will be described below. The fourth embodiment has the same circuit configuration as that of the first embodiment shown in FIG. During the heating operation, depending on the air conditions, frost formation occurs in the air heat exchanger 5, so it is necessary to perform a defrost operation. In general, when the defrost operation is performed, the heating operation is stopped, the heat supply to the load side is stopped, the defrost operation is performed, and the heating operation is resumed after completion. Therefore, during the defrost operation, the heat load cannot be covered, which may cause a problem in the comfort of the air conditioning operation.
本実施の形態では、デフロスト運転は、冷却運転と同様の運転動作を行い、圧縮機3からの過熱ガスを空気熱交換器5に供給することでデフロストを行う。このデフロスト運転を冷凍サイクル2a、2b交互に行うことで、加熱運転を停止せず、加熱運転を行いながらデフロスト運転を実施する。すなわち、冷凍サイクル2aがデフロスト運転を行う場合は、冷凍サイクル2bの方を加熱運転とし、冷凍サイクル2bから供給される温水を熱源に冷凍サイクル2aのデフロスト運転を行う。逆に、冷凍サイクル2bのデフロスト運転を行う場合は、冷凍サイクル2aの方を加熱運転とし、負荷側から供給される温水を熱源に冷凍サイクル2bのデフロスト運転を行う。   In the present embodiment, the defrost operation is performed by performing the same operation as the cooling operation and supplying the superheated gas from the compressor 3 to the air heat exchanger 5. By performing this defrost operation alternately with the refrigeration cycles 2a and 2b, the defrost operation is performed while the heating operation is performed without stopping the heating operation. That is, when the refrigeration cycle 2a performs the defrost operation, the refrigeration cycle 2b is set as a heating operation, and the defrost operation of the refrigeration cycle 2a is performed using hot water supplied from the refrigeration cycle 2b as a heat source. Conversely, when performing the defrost operation of the refrigeration cycle 2b, the refrigeration cycle 2a is set as a heating operation, and the defrost operation of the refrigeration cycle 2b is performed using hot water supplied from the load side as a heat source.
デフロスト運転実施時は、空気熱交換器5のファン11を停止する必要がある。従って上記のような片側デフロスト運転を行う場合は、各冷凍サイクル2の空気熱交換器5が独立した構成であるとともに、ファン11も各冷凍サイクル個別に配置する。このような配置とすることで、デフロスト運転が必要となる運転条件であっても加熱運転を停止することなく、連続して実施でき、快適性の高い空調装置の運転を実現できる。   When performing the defrost operation, it is necessary to stop the fan 11 of the air heat exchanger 5. Accordingly, when performing the one-side defrost operation as described above, the air heat exchanger 5 of each refrigeration cycle 2 has an independent configuration, and the fans 11 are also arranged individually for each refrigeration cycle. By adopting such an arrangement, even if the operating conditions require defrost operation, the heating operation can be continuously performed without stopping, and a highly comfortable air conditioner operation can be realized.
この発明の実施の形態1を示す冷凍空調装置の回路図である。1 is a circuit diagram of a refrigerating and air-conditioning apparatus showing Embodiment 1 of the present invention. この発明の実施の形態1に係わる冷凍空調装置の圧力とエンタルピの相関を示す図である。It is a figure which shows the correlation of the pressure and enthalpy of the refrigerating air-conditioning apparatus concerning Embodiment 1 of this invention. この発明の実施の形態1に係わる冷却運転での冷凍空調装置の制御動作を示す図である。It is a figure which shows the control action of the refrigerating air conditioner in the cooling operation concerning Embodiment 1 of this invention. この発明の実施の形態1に係わる加熱運転での冷凍空調装置の制御動作を示す図である。It is a figure which shows the control action of the refrigerating air-conditioning apparatus in the heating operation concerning Embodiment 1 of this invention. この発明の実施の形態1に係わる冷却運転時の熱源機の制御特性を示す図である。It is a figure which shows the control characteristic of the heat-source equipment at the time of the cooling operation concerning Embodiment 1 of this invention. この発明の実施の形態1に係わる冷却運転時の水熱交換器の温度差の特性を示す図である。It is a figure which shows the characteristic of the temperature difference of the water heat exchanger at the time of the cooling operation concerning Embodiment 1 of this invention. この発明の実施の形態1に係わる冷却運転時の水熱交換器での温度変化を示す図である。It is a figure which shows the temperature change in the water heat exchanger at the time of the cooling operation concerning Embodiment 1 of this invention. この発明の実施の形態2を示す冷凍空調装置の回路図である。It is a circuit diagram of the refrigerating and air-conditioning apparatus which shows Embodiment 2 of this invention. この発明の実施の形態3を示す冷凍空調装置の回路図である。It is a circuit diagram of the refrigerating and air-conditioning apparatus which shows Embodiment 3 of this invention.
符号の説明Explanation of symbols
1 熱源機、2a、2b 冷凍サイクル、3a、3b、3c 圧縮機、4a、4b 四方弁、5a、5b 空気熱交換器、6a、6b、6c、6d、6e、6f、6g、6h 逆止弁、7a、7b 過冷却熱交換器、8a、8b 主膨張弁、9a、9b 水熱交換器、10a、10b バイパス膨張弁、11a、11b ファン、12 ポンプ、13 計測制御装置、14a、14b、14c、14d 圧力センサ、15a、15b、15c、15d、15e、15f、15g、15h、15i、15j、15k、15l、15m、15n、15o、15p、15q、15r、15s 温度センサ。   1 heat source machine, 2a, 2b refrigeration cycle, 3a, 3b, 3c compressor, 4a, 4b four-way valve, 5a, 5b air heat exchanger, 6a, 6b, 6c, 6d, 6e, 6f, 6g, 6h check valve 7a, 7b Supercooling heat exchanger, 8a, 8b Main expansion valve, 9a, 9b Water heat exchanger, 10a, 10b Bypass expansion valve, 11a, 11b Fan, 12 Pump, 13 Measurement control device, 14a, 14b, 14c 14d Pressure sensor, 15a, 15b, 15c, 15d, 15e, 15f, 15g, 15h, 15i, 15j, 15k, 15l, 15m, 15n, 15o, 15p, 15q, 15r, 15s Temperature sensor.

Claims (8)

  1. 運転容量が可変である圧縮機と、熱源側熱交換器と、減圧装置と、負荷側熱交換器とを環状に接続して構成される冷凍サイクルを複数備え、各冷凍サイクルの負荷側熱交換器において負荷側熱媒体を冷却又は加熱し、冷温熱を供給するとともに、負荷側熱媒体の流路が各冷凍サイクルの負荷側熱交換器を直列に流れるように構成され、
    各冷凍サイクルの負荷側熱交換器における負荷側熱媒体の流入温度及び流出温度をそれぞれ計測する温度センサと、
    負荷側熱交換器が負荷側熱媒体の流路の下流側に接続される冷凍サイクルの圧縮機容量を、各冷凍サイクルの圧縮機運転容量が負荷側熱交換器の容量又は圧縮機の定格容量に比例するようにして決定される圧縮機運転容量よりも大きく、かつ、負荷側熱交換器が負荷側熱媒体の流路の下流側に接続される冷凍サイクルの負荷側熱媒体の流入、流出の温度差が、各冷凍サイクルに搭載される負荷側熱交換器の容量又は圧縮機の定格容量に比例するようにして決定される温度差よりも小さくなるように、圧縮機運転容量を制御する制御装置と
    を備えることを特徴とする冷凍空調装置。
    It is equipped with multiple refrigeration cycles configured by connecting a compressor with a variable operating capacity, a heat source side heat exchanger, a pressure reducing device, and a load side heat exchanger, and load side heat exchange of each refrigeration cycle The load-side heat medium is cooled or heated in the refrigerator to supply cold / hot heat, and the flow path of the load-side heat medium is configured to flow in series through the load-side heat exchanger of each refrigeration cycle,
    A temperature sensor for measuring the inflow temperature and the outflow temperature of the load-side heat medium in the load-side heat exchanger of each refrigeration cycle, and
    The compressor capacity of the refrigeration cycle in which the load side heat exchanger is connected to the downstream side of the flow path of the load side heat medium, the compressor operating capacity of each refrigeration cycle is the capacity of the load side heat exchanger or the rated capacity of the compressor Inflow and outflow of the load-side heat medium of the refrigeration cycle that is larger than the compressor operating capacity determined in proportion to the load and the load-side heat exchanger is connected to the downstream side of the flow path of the load-side heat medium The compressor operating capacity is controlled so that the temperature difference is smaller than the temperature difference determined in proportion to the capacity of the load-side heat exchanger installed in each refrigeration cycle or the rated capacity of the compressor. A refrigeration air conditioner comprising a control device.
  2. 前記制御装置は、各冷凍サイクルの圧縮機を起動する場合には、負荷側熱交換器が負荷側熱媒体の流路の下流側に接続される冷凍サイクルの圧縮機を先に起動するように制御することを特徴とする請求項1に記載の冷凍空調装置。   When starting the compressor of each refrigeration cycle, the control device starts the refrigeration cycle compressor connected to the downstream side of the flow path of the load-side heat medium first by the load-side heat exchanger. The refrigerating and air-conditioning apparatus according to claim 1, wherein the refrigerating and air-conditioning apparatus is controlled.
  3. 冷凍サイクルに搭載される圧縮機の定格容量が大きい冷凍サイクルの負荷側熱交換器が、冷凍サイクルに搭載される圧縮機の定格容量が小さい冷凍サイクルの負荷側熱交換器よりも、負荷側熱媒体の流路の上流に配置されることを特徴とする請求項1又は2に記載の冷凍空調装置。   The load-side heat exchanger of the refrigeration cycle with a large rated capacity of the compressor installed in the refrigeration cycle has a higher load-side heat than the load-side heat exchanger of the refrigeration cycle with a small rated capacity of the compressor installed in the refrigeration cycle. The refrigerating and air-conditioning apparatus according to claim 1, wherein the refrigerating and air-conditioning apparatus is disposed upstream of a flow path of the medium.
  4. 負荷側熱交換器を流出する冷媒をさらに冷却する過冷却熱交換器を設けたことを特徴とする請求項1〜3の何れか一項に記載の冷凍空調装置。   The refrigeration air conditioner according to any one of claims 1 to 3, further comprising a supercooling heat exchanger that further cools the refrigerant flowing out of the load side heat exchanger.
  5. 負荷側熱交換器を流出した冷媒を一部分岐しバイパスするバイパス回路を備え、バイパス回路上に分岐された冷媒を減圧する第二の減圧装置と、第二の減圧装置を流出した冷媒と負荷側熱交換器を流出した冷媒を熱交換する過冷却熱交換器とを備えるとともに、圧縮機にガスインジェクションが行われるポートを備え、過冷却熱交換器を流出したバイパス回路の冷媒が前記圧縮機ポートに流入するようにバイパス回路を構成することを特徴とする請求項4に記載の冷凍空調装置。   A second pressure reducing device for depressurizing the refrigerant branched on the bypass circuit, a second pressure reducing device for depressurizing the refrigerant branched on the bypass circuit, and the load side And a supercooling heat exchanger for exchanging heat from the refrigerant that has flowed out of the heat exchanger, a compressor is provided with a port for gas injection, and the refrigerant in the bypass circuit that has flowed out of the supercooling heat exchanger is connected to the compressor port The refrigerating and air-conditioning apparatus according to claim 4, wherein a bypass circuit is configured to flow into the refrigeration unit.
  6. 過冷却熱交換器において、負荷側熱交換器を流出する冷媒と圧縮機に吸入される冷媒が熱交換する構成とすることを特徴とする請求項4に記載の冷凍空調装置。   The refrigerating and air-conditioning apparatus according to claim 4, wherein in the subcooling heat exchanger, heat is exchanged between the refrigerant flowing out of the load side heat exchanger and the refrigerant sucked into the compressor.
  7. 熱源側熱交換器が空気と熱交換する構成とするとともに、空気を送風するファンが各冷凍サイクルに個別に設けられることを特徴とする請求項1〜6の何れか一項に記載の冷凍空調装置。   The refrigeration air conditioning system according to any one of claims 1 to 6, wherein the heat source side heat exchanger is configured to exchange heat with air, and a fan that blows air is individually provided in each refrigeration cycle. apparatus.
  8. 負荷側熱媒体を加熱する運転モードにおいて、負荷側熱交換器の流路を、冷媒と負荷側熱媒体とが対向して流れる流路構成とすることを特徴とする請求項1〜7の何れか一項に記載の冷凍空調装置。   The operation mode for heating the load-side heat medium, wherein the flow path of the load-side heat exchanger has a flow path configuration in which the refrigerant and the load-side heat medium flow oppositely. The refrigeration air conditioner according to claim 1.
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