JPS60140048A - Refrigerator using non-eutectic mixed refrigerant - Google Patents
Refrigerator using non-eutectic mixed refrigerantInfo
- Publication number
- JPS60140048A JPS60140048A JP24669083A JP24669083A JPS60140048A JP S60140048 A JPS60140048 A JP S60140048A JP 24669083 A JP24669083 A JP 24669083A JP 24669083 A JP24669083 A JP 24669083A JP S60140048 A JPS60140048 A JP S60140048A
- Authority
- JP
- Japan
- Prior art keywords
- heat exchanger
- mixed refrigerant
- refrigerant
- temperature
- pressure
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Pending
Links
Landscapes
- Compression-Type Refrigeration Machines With Reversible Cycles (AREA)
Abstract
(57)【要約】本公報は電子出願前の出願データであるた
め要約のデータは記録されません。(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.
Description
（技術分野〕
本発明は非共沸混合冷媒を用いた冷凍装置、詳しくは圧
縮機、熱源側熱交換器、利用側熱交換器及び膨張機構を
備えた冷媒回路に非共沸混合冷媒を封入し、前記利用側
熱交換器を凝縮器とじて暖房可能とした冷凍装置に関す
る。
（従来波＃ｉ］
上記した如く非共沸混合冷媒を封入した冷凍装置は特開
昭５７−１９８９６８号公報にも記載されずでに知られ
ている。
ところが、上記従来のものは、暖房運転時に蒸発器とし
て作用する熱源側熱交換器において冷媒を一定の圧力で
蒸発させていたために、前記熱源側熱交換器の能力が十
分発揮させられず、装置全体の性能が低下する問題があ
った。
以下、この点を具体的に説明する。
非共沸混合冷媒を、一定圧力で蒸発させていくと、単一
冷媒と異なり蒸発器こ従って液冷媒の組成が変化し、こ
れに伴い蒸発温度が上昇していくのである。（以下、こ
れを温度勾配とい′）ｃ、）この点を＠４図モリエル線
で示すと、蒸発器として作用する熱交換器ｃｐＸＪ示せ
ず】での混合冷媒の蒸発圧力Ｐは従来一定である力）ら
、前記熱交換器の入口側（イ］での蒸発温度がＴ＋、出
口側（ロ）でのそれが’ｆｌより高い温度Ｔｗとなり、
入Ｏａと出口側とで大きな蒸発温度差を生じるのである
。この温度差（ΔＴ ＝ Ｔ＊ −’Ｌ’ｌ ）は封入
する混合冷媒の組成によって異なるが１例えば１５℃に
も及ぶのである。尚、第４図１点＠線で示したものは等
温線である。このため、温度勾配の大きい混合冷媒を用
いると、熱源側熱交換器での最適な蒸発温度域が、外気
温度（標準７°Ｃ）とフロスト限界温度（約−６°Ｃ）
とにより狭く限定される暖房運転時５こおいては１例え
ば、
■ 蒸発圧力を高めに設定して、前記熱交換器の人口で
の蒸発温度を高くシ、フロストの問題を回避しようとす
ると、前記熱交換器の出口側では蒸発温度が外気温度よ
り高くなり、該熱交換器が作用できない無効域が増大す
るし、又、■ 逆に、蒸発圧力を低めに設定して前記熱
交換器の出口側における前記無効域を無くそうとすると
、今度は該熱交換器の人口側での蒸発温度がフロスト限
界温度以下となりフロストの問題を生じるのであって、
いずれにしても、前記熱交換器の能力が十分発揮できず
、このため装置全体の能力が低下する問題があったので
ある。
〔発明の目的〕
本発明は暖房運転時に、前記熱源側熱交換器での蒸発圧
力を、出口側に行くに従って低下するように変化させて
やれば、前記熱交換器内での蒸発温度の上昇幅を小さく
できる点に着目して発明したもので、目的は、暖房運転
特において熱源側熱交換器における７０ストの問題なら
びに前記無効域の発生の問題を共に解決できるようにし
て。
装置全体の性能を向上させる点にある。
（発明の構成）
而して、本発明の構成は圧縮機、熱源側熱交換器、利用
側熱交換器及び膨張機構を備えた冷媒回路に非共沸混合
冷媒を封入し、前記利用側熱交換器を凝縮器として暖房
可能とした冷凍装置において、暖房運転時蒸発器となる
前記熱源側熱交換器に、該熱交換器を流れる前記混合冷
媒を、該混合冷媒の温度勾配に対応して減圧する減圧手
段を設けて、暖房運転時、前記熱源側熱交換器における
冷媒の蒸発圧力を出口側に向って低下させることにより
、蒸発温度の上昇幅を小さくできるように成したのであ
る。
（実 施 例）
以下、本発明の一実施例を図面に基づいて説明する。
′＠１図に示したものは、高沸点冷媒Ｒ−２２と低沸点
冷媒Ｒ−１６とを混合した非共沸混合冷媒を封入するヒ
ートポンプ式の冷凍装置である。
図中、（１］は圧縮機、（２）は利用側熱交換器、（６
）は熱源側熱交換器、（４］は膨張機構として作用する
第１キヤピラリーチユーブであり、これら機器は四路切
換弁（５］を介して可逆サイクルを構成する如く接続さ
れている。尚、（６］はアキュムレータである。
而して１以上の始〈構成する冷凍装置において、前記熱
源側熱交換器（３）を、第１熱交換器（３１）と第２熱
交換器〔６２〕とに分割し、かつ、これら熱交換器（３
１）、（３２）を減圧手段メなる第２キヤピラリーチユ
ーブ（７）を介して直列に接続して構成し、このことに
より暖房運転時に、前記混合冷媒の蒸発圧力をこの熱源
側熱交換器(Technical field) The present invention relates to a refrigeration system using a non-azeotropic mixed refrigerant, specifically, a refrigerant circuit equipped with a compressor, a heat source side heat exchanger, a user side heat exchanger, and an expansion mechanism, in which the non-azeotropic mixed refrigerant is sealed. The invention relates to a refrigeration system in which heating is possible by combining the heat exchanger on the user side with a condenser. (Conventional wave #i) A refrigeration system in which a non-azeotropic mixed refrigerant is sealed as described above is disclosed in Japanese Patent Application Laid-open No. 198968/1983. However, in the above conventional system, the refrigerant is evaporated at a constant pressure in the heat source side heat exchanger that acts as an evaporator during heating operation. There was a problem that the capacity of the refrigerant was not fully utilized, and the performance of the entire equipment decreased.This point will be explained in detail below.When a non-azeotropic mixed refrigerant is evaporated at a constant pressure, Unlike a single refrigerant, the composition of the liquid refrigerant changes in the evaporator, and the evaporation temperature rises accordingly. Since the evaporation pressure P of the mixed refrigerant in the heat exchanger cpXJ, which acts as an evaporator (not shown), is conventionally constant, the evaporation temperature at the inlet side (a) of the heat exchanger is T+, At the exit side (b), it becomes a temperature Tw higher than 'fl,
This creates a large evaporation temperature difference between the inlet Oa and the outlet side. This temperature difference (ΔT = T* -'L'l) varies depending on the composition of the mixed refrigerant to be sealed, but can be as high as 15°C, for example. Note that the one point @ line in FIG. 4 is an isothermal line. Therefore, when using a mixed refrigerant with a large temperature gradient, the optimal evaporation temperature range in the heat source side heat exchanger is between the outside air temperature (standard 7°C) and the frost limit temperature (approximately -6°C).
During heating operation, which is narrowly limited by On the outlet side of the heat exchanger, the evaporation temperature becomes higher than the outside air temperature, increasing the ineffective area where the heat exchanger cannot function; If an attempt is made to eliminate the ineffective area on the outlet side, the evaporation temperature on the population side of the heat exchanger will be below the frost limit temperature, causing a frost problem. There was a problem in that the capacity could not be fully demonstrated, and as a result, the capacity of the entire device decreased. [Object of the Invention] The present invention provides that during heating operation, if the evaporation pressure in the heat source side heat exchanger is changed so as to decrease toward the outlet side, the evaporation temperature in the heat exchanger can be increased. This invention was developed focusing on the ability to reduce the width, and the purpose was to solve both the problem of 70 strokes in the heat source side heat exchanger and the problem of the generation of the ineffective region, especially during heating operation. The purpose is to improve the performance of the entire device. (Structure of the Invention) Therefore, the structure of the present invention is that a non-azeotropic mixed refrigerant is sealed in a refrigerant circuit equipped with a compressor, a heat source side heat exchanger, a user side heat exchanger, and an expansion mechanism, and In a refrigeration system in which heating is possible using an exchanger as a condenser, the mixed refrigerant flowing through the heat exchanger is supplied to the heat source side heat exchanger that serves as an evaporator during heating operation in a manner corresponding to the temperature gradient of the mixed refrigerant. A pressure reducing means is provided to reduce the evaporation pressure of the refrigerant in the heat source side heat exchanger toward the outlet side during heating operation, thereby making it possible to reduce the rise in evaporation temperature. (Example) Hereinafter, one example of the present invention will be described based on the drawings. '@1 What is shown in Fig. 1 is a heat pump type refrigeration system in which a non-azeotropic mixed refrigerant, which is a mixture of a high boiling point refrigerant R-22 and a low boiling point refrigerant R-16, is sealed. In the figure, (1) is the compressor, (2) is the user-side heat exchanger, (6
) is a heat source side heat exchanger, and (4) is a first capillary reach tube that acts as an expansion mechanism, and these devices are connected to form a reversible cycle via a four-way switching valve (5). , (6) is an accumulator.In the refrigeration system comprising one or more sources, the heat source side heat exchanger (3) is connected to the first heat exchanger (31) and the second heat exchanger [62]. ] and these heat exchangers (3
1) and (32) are connected in series through a second capillary reach tube (7) serving as a pressure reducing means, thereby controlling the evaporation pressure of the mixed refrigerant to the heat source side heat exchanger during heating operation.
【３】の流通過程で減圧できるように成すの
である。
尚、（８〕は、冷房運転時前記混合冷媒を前記第２キヤ
ピラリーチユーブ〔７〕をバイパスさせて循環させるた
めの逆止弁である。
更に、前記各キャピラリーチューブ（４］ｌ（７］によ
る各減圧量の設定について、第３図のモリエル線図を基
に説明する。
尚第５図中、Ｔｉは７ｐスト限界温度（例えば−３°Ｃ
）、Ｔ！は暖房運転時における標準外気温度（例えば７
°Ｃ）を示す等温線である。
前記第１熱交換器（３１）の入口側にある前記第１キヤ
ピラリーチユーブ（７）の減圧量は、暖房運転時前記熱
交換器（３１］の人口での混合冷媒の蒸発温度が前記７
０スト限界温度１より僅７３）に高い温度ＴＩとなる圧
力ＰＫに成るように設定する。
又、第１．２熱交換器（３１）、（３２）間に設ける前
記第２キヤピラリーチユーブ（７〕の減圧量は前記混合
冷媒の温度勾配に対応させて決定するもので、具体的に
は、前記第２熱交換器（３２）の人口での蒸発温度が前
記フロスト限界温度ＴＩ以上の温度Ｔ＠で、かつ出口で
の蒸発温度が前記標準外気温度（Ｔｆｉ）を適当に下回
る温度Ｔ・となる圧力Ｐ１まで減圧する如く設定すＱの
である。
次に以上の如く構成する冷凍装置の作用を説明する。
暖房運転について説明する。前記四路切換弁（５）を第
１図実線で示す如く切換えて暖房サイクルを形成する。
そして、圧縮機（１〕を駆動させると、混合冷媒は実線
矢印で示す１ｘＪ（循環する。この循環による混合冷媒
の状態変化を＠１図および第６図のモリエル線図を基に
説明する。
混合冷媒は圧縮機（１〕力）ら圧力Ｐｏの高温高圧のガ
スとして吐出されるし）。そして利用側熱交換器（２］
で同一圧力（ＰＯ３で凝縮して液冷媒ドア２ル（二）。
更≦こ、前記第１キヤピラリーチユーブ（４］で減圧さ
れて圧力Ｐ１の状態となって前記第１熱交換器（６１）
に流入する（ホ）。
而して前記混合冷媒は前記第１熱交換器（３１）の人口
で７０スト限界温度〕より高い温度Ｔｓで蒸発を開始し
、この蒸発に伴い該熱交換器□Ｓ ３１】の出口側では
蒸発温度がＴｉ４（但しＴ雪量下）まで上昇するのであ
る（へ〕。そして、この第１熱交換器（３１）から流出
した前記混合冷媒は前記第２キヤピラリーチユーブ（７
）で再び減圧されて圧力ｐｇとなるのであり（ト）、こ
のため前記第２熱交換器〔３２〕の人口での蒸発温度は
前記第１熱交換器（３１）の出口側のそれより低く、し
力）もフロスト限界温度ＴＩより高い温度Ｔ１に再び低
下するのであるThis is done so that the pressure can be reduced during the distribution process (3). In addition, (8) is a check valve for circulating the mixed refrigerant while bypassing the second capillary reach tube [7] during cooling operation. Furthermore, each of the capillary tubes (4) l (7) The setting of each pressure reduction amount will be explained based on the Mollier diagram shown in Fig. 3. In Fig. 5, Ti is set at 7 ps strike limit temperature (for example -3°C
), T! is the standard outside temperature during heating operation (e.g. 7
℃). The amount of pressure reduction in the first capillary reach tube (7) located on the inlet side of the first heat exchanger (31) is such that the evaporation temperature of the mixed refrigerant at the population of the heat exchanger (31) during heating operation is 7.
The pressure PK is set so that the temperature TI is only 73) higher than the zero stroke limit temperature 1. Further, the amount of pressure reduction in the second capillary reach tube (7) provided between the first and second heat exchangers (31) and (32) is determined in accordance with the temperature gradient of the mixed refrigerant. is a temperature T at which the evaporation temperature at the population of the second heat exchanger (32) is higher than the frost limit temperature TI, and the evaporation temperature at the outlet is appropriately lower than the standard outside air temperature (Tfi). Q is set so as to reduce the pressure to a pressure P1 of The switching is performed as shown to form a heating cycle. Then, when the compressor (1) is driven, the mixed refrigerant circulates 1xJ (as shown by the solid arrow). The changes in the state of the mixed refrigerant due to this circulation are shown in Figure 1 and Figure 6. This will be explained based on the Mollier diagram.The mixed refrigerant is discharged from the compressor (1) as a high temperature, high pressure gas with a pressure Po).Then, the user side heat exchanger (2)
At the same pressure (PO3 is condensed and the liquid refrigerant is refrigerated at the door 2 (2). )
(e). Therefore, the mixed refrigerant starts to evaporate at a temperature Ts higher than the 70 stroke limit temperature of the first heat exchanger (31), and with this evaporation, on the outlet side of the heat exchanger The evaporation temperature rises to Ti4 (but below the T snow amount).Then, the mixed refrigerant flowing out from the first heat exchanger (31) flows into the second capillary reach tube (7).
), and the pressure is reduced to pg (g). Therefore, the evaporation temperature at the second heat exchanger [32] is lower than that at the outlet side of the first heat exchanger (31). , force) decreases again to a temperature T1 higher than the frost limit temperature TI.
【ト）。更に、この第２熱交換器（６２
）での蒸発に伴い蒸発温度が上昇し、出口側において、
標準外気温度Ｔｓより適当に低い温度Ｔ６のガス冷媒と
なって（チ）、前記圧縮機（１］に返送され再び圧縮さ
れるのである。
以上の叩り、前記熱源側熱交換器〔６〕に減圧手段（第
２キヤピラリーチユーブ）（７）を設けた力）ら、前記
熱交換器〔６〕の入口側と出口側との蒸発温度の差、換
言すると蒸発温度の上昇幅が小さくなり、前記蒸発温度
を前記した最適蒸発温度域内とすることができるのであ
って、この結果、前記熱交換器（３］でのフロストを回
避、又は抑制できながら、しかも前記無効域も生じるこ
とがないのである。
更に、上記の如く、前記温度勾配の大きい混合冷媒を使
用しても、前記熱源側熱交換器（３）の能力低下をきた
さないから、冷凍装置に使用する混合冷媒の組成の選択
幅も拡がり、使用状態に則した最適のものが選べるので
ある。
尚、本実施例は、前記四路切換弁（５）を破線で示す如
く切換え、破線矢印で示すサイクルを形成して冷房運転
を行えるものであるが、これは従来と同様であるから説
明を省略する。
次に本発明の第２実施例を第２図に基づいて説明する。
この実施例も第１実施例と同様に非共沸点冷媒を対人し
たヒートポンプ式の冷凍装置である。
第１実施例との相違点は、負荷に応じて、前記混合冷媒
の組成を変えて能力を増減させられるようにしている点
である。具体的には、膨張機構として作用する第３．第
４キヤピラリーチユーブ（４１）、（４２）の間に気液
分離器（９〕を設ける一方、吸入ガス管（１０）に冷媒
貯留用の容器（１１）を相互に熱交換可能に付設し、該
容器（１１）の一端を前記分離器（９］のガス域に。
他端を吸入ガス管（１０）に接続し、これら接続回路に
それぞれ第１．第２開閉弁（１２）、（１３）を設ける
如く成している。
そして、前記第２開閉弁（１３）を閉にし、かつ第１開
閉弁（１２）を開にすることにより、前記容器【１１】
に前記分離器（９〕力）ら低沸点冷媒の多い混合冷媒を
流入させて、凝縮貯留させられるのである。このことに
より循環する混合冷媒における高沸点冷媒の組成比が大
きくなって能力が低下するのである。
また、前記事１．第２開閉弁（１２）、（１６〕を前記
したのと逆に開閉操作すると、再び混合冷媒の組成比が
元に戻り能力がアップするのである。
その他の構成は第１実施例と同様であるから、第２図に
第１実施例の構成と同一符号を付して説明を省略する。
尚、上記実施例においては、前記熱源側熱交換器（６）
における蒸発圧力を２ステツプとしたが、前記熱交換器
（３）を３個以上に分割して設け、これら分割した各熱
交換器量器こそれぞれキャピラリーチューブを介装して
ろステップ以上に蒸発圧力を変化させてもよい。
また、上記実施例においては、前記熱源側熱交換器２（
３）に減圧手段としてキャピラリーチューブ（７）を設
けたが、前記熱交換器（３）に設ける混合冷媒流通用の
バスの内径を、適当な減圧勾配が得られるように選定す
ることにより前記減圧手段を構成してもよい。
更に、前記減圧手段の減圧量は、必ずしも暖房運転時に
おける前記熱源側熱交換器（６］の入口での蒸発温度が
前記フロスト限界温度以上となるようにしなくともよい
。
（発明の効果］
以上の如く、本発明は暖房運転時蒸発器となる前記熱源
側熱交換器（３）に、該熱交換器（３］を流れる前記混
合冷媒を、該混合冷媒の温度勾配に対応して減圧する減
圧手段（７）を設けたから、使用する混合冷媒の温度勾
配が大きくとも暖房運転特における熱源側熱交換器（３
）における７０ストの問題を回避もしくは抑制できなが
ら。
しかも前記無効域の発生も防止して前記熱交換器（６〕
の能力が有効に利用でき、この結果、装置全体の性能の
向上が計れるのである。【to). Furthermore, this second heat exchanger (62
), the evaporation temperature increases, and at the outlet side,
The gas refrigerant becomes a gas refrigerant at a temperature T6 suitably lower than the standard outside air temperature Ts (H), and is returned to the compressor (1) and compressed again. After the above beating, the heat source side heat exchanger [6] The difference in evaporation temperature between the inlet side and the outlet side of the heat exchanger [6], in other words, the increase in the evaporation temperature becomes smaller due to the pressure reduction means (second capillary reach tube) (7) provided in the heat exchanger [6]. , the evaporation temperature can be set within the optimum evaporation temperature range described above, and as a result, frost in the heat exchanger (3) can be avoided or suppressed, and the ineffective region does not occur. Furthermore, as mentioned above, even if the mixed refrigerant with a large temperature gradient is used, the capacity of the heat source side heat exchanger (3) will not be reduced, so the selection of the composition of the mixed refrigerant used in the refrigeration system is The range is widened, and the most suitable one can be selected according to the usage condition.In this embodiment, the four-way switching valve (5) is switched as shown by the broken line, and a cycle shown by the broken line arrow is formed to perform cooling operation. This is the same as the conventional method, so the explanation will be omitted. Next, a second embodiment of the present invention will be explained based on FIG. 2. This embodiment is also similar to the first embodiment. This is a heat pump type refrigeration system using a non-azeotropic refrigerant.The difference from the first embodiment is that the capacity can be increased or decreased by changing the composition of the mixed refrigerant depending on the load. Specifically, a gas-liquid separator (9) is provided between the third and fourth capillary reach tubes (41) and (42) that act as an expansion mechanism, while a gas-liquid separator (9) is provided in the suction gas pipe (10) for refrigerant storage. containers (11) are attached so as to be able to exchange heat with each other, one end of the container (11) is connected to the gas region of the separator (9), the other end is connected to the suction gas pipe (10), and these connecting circuits are connected. First and second on-off valves (12) and (13) are respectively provided in the two on-off valves (12) and (13).Then, the second on-off valve (13) is closed and the first on-off valve (12) is opened. Possibly, the container [11]
A mixed refrigerant containing a large amount of low boiling point refrigerant is allowed to flow into the separator (9) and is condensed and stored. This increases the composition ratio of the high boiling point refrigerant in the circulating mixed refrigerant, resulting in a decrease in performance. Also, see the previous article 1. When the second on-off valves (12) and (16) are opened and closed in the opposite manner to the above, the composition ratio of the mixed refrigerant returns to its original state and the capacity increases.Other configurations are the same as in the first embodiment. Therefore, the same reference numerals as those in the configuration of the first embodiment are given in FIG. 2, and the explanation thereof will be omitted.
The heat exchanger (3) is divided into three or more units, and each of these divided heat exchanger meters is interposed with a capillary tube to increase the evaporation pressure to a level higher than the two steps. It may be changed. Further, in the above embodiment, the heat source side heat exchanger 2 (
3), a capillary tube (7) is provided as a pressure reduction means, but the pressure reduction can be achieved by selecting the inner diameter of the mixed refrigerant circulation bath provided in the heat exchanger (3) so as to obtain an appropriate pressure reduction gradient. The means may also be configured. Furthermore, the amount of pressure reduction by the pressure reduction means does not necessarily have to be such that the evaporation temperature at the inlet of the heat source side heat exchanger (6) during heating operation becomes equal to or higher than the frost limit temperature. (Effects of the Invention) As shown in the figure, the present invention reduces the pressure of the mixed refrigerant flowing through the heat exchanger (3) to the heat source side heat exchanger (3), which serves as an evaporator during heating operation, in accordance with the temperature gradient of the mixed refrigerant. Since the pressure reducing means (7) is provided, even if the temperature gradient of the mixed refrigerant used is large, the heat source side heat exchanger (3) is
) while avoiding or suppressing the 70-stroke problem. Moreover, the generation of the ineffective area is also prevented, and the heat exchanger (6)
can be used effectively, and as a result, the performance of the entire device can be improved.
第１図は本発明の第１実施例を示す冷媒回路図、′＠２
図は他の実施例を示す冷媒回路図％第６図は第１実施例
の運転状態を説明するための説明図、第４図は従来例を
説明するための説明図である。
（１）・・・圧縮機
（２）・・・利用側熱交換器
（３）・・・熱源側熱交換器
（７）・・・第２キヤピラリーチユーブ（減圧手段］
符閏昭ＧＯ−１４００４８（５）
エンタノＵと０−FIG. 1 is a refrigerant circuit diagram showing a first embodiment of the present invention, '@2
FIG. 6 is a refrigerant circuit diagram showing another embodiment. FIG. 6 is an explanatory diagram for explaining the operating state of the first embodiment, and FIG. 4 is an explanatory diagram for explaining the conventional example. (1) Compressor (2) Utilization side heat exchanger (3) Heat source side heat exchanger (7) Second capillary reach tube (pressure reduction means) 5) Entano U and 0-
Claims (1)
熱交換器（２）及び膨張機構（４）を備えた冷媒回路に
非共沸混合冷媒を封入し、前記利用側熱交換器（２〕を
凝縮器として暖房可能とした冷凍装置において、暖房運
転時蒸発器となる前記熱源側熱交換器（３）に、該熱交
換器（５〕を流れる前記混合冷媒を、該混合冷媒の温度
勾配に対応して減圧する減圧手段（７）を設けたことを
特徴とする非共沸混合冷媒を用いた冷凍装置。(1) A non-azeotropic mixed refrigerant is sealed in a refrigerant circuit equipped with a compressor [1], a heat source side heat exchanger (3), a user side heat exchanger (2), and an expansion mechanism (4), and In a refrigeration system capable of heating using a heat exchanger (2) as a condenser, the mixed refrigerant flowing through the heat exchanger (5) is transferred to the heat source side heat exchanger (3) which becomes an evaporator during heating operation. A refrigeration system using a non-azeotropic mixed refrigerant, characterized in that a pressure reducing means (7) is provided to reduce the pressure in accordance with the temperature gradient of the mixed refrigerant.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP24669083A JPS60140048A (en) | 1983-12-27 | 1983-12-27 | Refrigerator using non-eutectic mixed refrigerant |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP24669083A JPS60140048A (en) | 1983-12-27 | 1983-12-27 | Refrigerator using non-eutectic mixed refrigerant |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| JPS60140048A true JPS60140048A (en) | 1985-07-24 |
Family
ID=17152174
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP24669083A Pending JPS60140048A (en) | 1983-12-27 | 1983-12-27 | Refrigerator using non-eutectic mixed refrigerant |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPS60140048A (en) |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH10339511A (en) * | 1997-06-09 | 1998-12-22 | Yoriyuki Oguri | Heat pump system of air conditioner |
| JP2010216778A (en) * | 2009-03-19 | 2010-09-30 | Hitachi Appliances Inc | Refrigerating cycle device |
| WO2021171401A1 (en) * | 2020-02-26 | 2021-09-02 | 三菱電機株式会社 | Air conditioning apparatus |
Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5248145A (en) * | 1975-10-14 | 1977-04-16 | Mitsubishi Electric Corp | Refrigerator with freezing compartment |
-
1983
- 1983-12-27 JP JP24669083A patent/JPS60140048A/en active Pending
Patent Citations (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5248145A (en) * | 1975-10-14 | 1977-04-16 | Mitsubishi Electric Corp | Refrigerator with freezing compartment |
Cited By (5)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH10339511A (en) * | 1997-06-09 | 1998-12-22 | Yoriyuki Oguri | Heat pump system of air conditioner |
| JP4208982B2 (en) * | 1997-06-09 | 2009-01-14 | グリーンアース株式会社 | Heat pump air conditioner |
| JP2010216778A (en) * | 2009-03-19 | 2010-09-30 | Hitachi Appliances Inc | Refrigerating cycle device |
| WO2021171401A1 (en) * | 2020-02-26 | 2021-09-02 | 三菱電機株式会社 | Air conditioning apparatus |
| JPWO2021171401A1 (en) * | 2020-02-26 | 2021-09-02 |
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