JPH06307738A - Condenser for non-azeotrope reefrigerant - Google Patents

Abstract

PURPOSE:To dissolve a provlem wherein at a freezing cycle using a non- azeotrope refrigerant, when a conventional condenser is used, complete liquefaction is impossible, the size of a comdenser is increased, energy efficiency is lowered, and the generation of refrigerant noise in an expansion valve is increased. CONSTITUTION:Refrigerant gas flows in through a refrigerant entrance 5 and liquefied refrigerant gas flows out through a refrigerant outlet 6. A condenser is a cross fin tube-type heatexchanger wherein a heat-transfer pipe 1 is extended through a fin 2. A high performance fin 3 is used in the vicinity of the refrigerant outlet 6, completion of condensation of a non-azeotrope refigerant is hastened, and the size of the condenser remains same as that of a conventional single refrigerant. As the result of condensation in the vicinity of the condensation completion point of the non-azeotrope refrigerant being promoted, the size of a condenser is same as that of a single refrigerant, lowering of COPis prevented, and the generation of refrigerant noise by an expansion valve is reduced effectively.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a condenser installed in an outdoor unit of an air conditioner exclusively for cooling, which uses a non-azeotropic mixed refrigerant obtained by mixing two or more kinds of refrigerants, or an air conditioner for both heating and cooling. The present invention relates to a heat exchanger installed indoors or outdoors.

[0002]

2. Description of the Related Art Regarding the condenser of the cross fin tube type heat exchanger used in the present room air conditioners and package air conditioners, there is no conventional example in which the non-azeotropic mixed refrigerant is specially devised.

However, regarding the plate type heat exchanger used in the binary power generation system, Japanese Patent Laid-Open No. 4-298.
There is a known example as shown in Japanese Patent No. 606. Japanese Patent Laid-Open No. Hei 4
In the plate heat exchanger described in JP-A-298606, a liquid having a high concentration of a high boiling point component is injected from a condenser inlet to promote the condensation of the low boiling point component gas. In such a large system, an additional function such as an injection device can be added, but it is difficult to add the injection device in the air conditioner targeted by the present invention.

Further, according to Japanese Patent Publication No. 4-29842, it is clarified that the components which are not completely condensed in the condenser are separated by the gas-liquid separator and this gas is returned to the inlet of the condenser again. There is. However, this is also a device related to the condenser in the power generation system, and it is difficult to apply this as it is to the condenser for the air conditioner.

[0005]

As described above, the conventional techniques disclosed in Japanese Patent Laid-Open No. 4-298606 and Japanese Patent Publication No. 4-29842 are difficult to apply to a condenser for an air conditioner. there were. The condenser for an air conditioner has a problem as described below regarding the condensation process of the non-azeotropic mixed refrigerant.

FIG. 6 is a diagram showing a vapor-liquid equilibrium diagram of a two-component system. A vapor having a certain molar component ratio starts to condense at a point A shown in FIG. 6 when the temperature decreases. However, the components of the condensate have passed through BD and have a high molar component ratio of high boiling point. On the other hand, the component remaining in the vapor passes through AC, so that the low boiling point molar component ratio is high. That is,
It can be seen that the low boiling point component that is difficult to condense becomes thicker as the condensation proceeds, and it becomes more difficult to condense.

Further, there is a temperature difference between the condensation start point A and the condensation completion point D, and the condensation temperature gradually decreases. Figure 7
FIG. 4 is a diagram showing changes in the refrigerant temperature, with the horizontal axis representing the refrigerant flow direction in the condenser and the vertical axis representing temperature. When the mixed refrigerant is condensed as shown by a in FIG. 7,
The temperature tends to decrease. On the other hand, with a single refrigerant, the same temperature is maintained until the condensation completion point as long as the pressure is constant. As a result, assuming that the cooling air temperature is c, the temperature difference between the mixed refrigerant and the cooling air is ΔTa at the condensation completion point, but the temperature difference between the single refrigerant and the cooling air is ΔTb. Since ΔTa is smaller than ΔTb, there is a first problem that the performance of the mixed refrigerant as a condenser is reduced and the required size of the condenser is increased.

The second problem is that, as shown in FIG. 8, the latent heat gradually increases as the condensation progresses. The latent heat of a single refrigerant is constant (indicated by a straight line b in FIG. 8) from the start of condensation to completion. On the other hand, in the mixed refrigerant, the condensed gas phase component changes like AC, and accordingly, the latent heat increases. The larger latent heat means that the condensation does not proceed unless more heat is taken. As described in the first problem, since the latent heat becomes large where the temperature difference is small, the size of the condenser still required becomes large.

The third problem is that, as shown in FIG. 9, the heat transfer coefficient sharply decreases near the condensation completion point. In particular, the heat transfer coefficient in the liquid region is extremely small, and the size of the condenser required for subcooling (hereinafter, also referred to as supercooling degree) is considerably large. Moreover, the low-boiling-point vapor remaining in the liquid does not easily disappear, and in many cases, it is a liquid containing bubbles forever. When this reaches the expansion valve as it is, it causes a large refrigerant noise.

As described above, the temperature difference is small,
Due to the three causes of large latent heat and small heat transfer coefficient, in the condenser of non-azeotropic mixed refrigerant, it is difficult to complete the condensation, and in order to complete the condensation and give a certain amount of subcool, a single refrigerant is required. It is estimated that the size of the condenser needs to be increased by about 10% compared with the above.

An object of the present invention is to provide a condenser for a non-azeotropic mixed refrigerant which can be applied to a condenser of a cross fin tube type heat exchanger even when a non-azeotropic mixed refrigerant is used.

[0012]

In order to achieve the above object, a condenser for a non-azeotropic mixed refrigerant of the present invention is a refrigeration cycle using a non-azeotropic mixed refrigerant in which two or more kinds of refrigerants are mixed. In the condenser, the fins on the downstream side of the refrigerant having a lower dryness than the fins on the upstream side of the refrigerant having a higher dryness are constituted by high-performance fins.

A condenser of a refrigeration cycle using a non-azeotropic mixed refrigerant in which two or more kinds of refrigerants are mixed, wherein the heat transfer tube on the upstream side of the refrigerant having a high degree of dryness has a downstream side of the refrigerant having a low degree of dryness. It is characterized in that the heat transfer tube of is composed of a heat transfer tube having a smaller diameter.

[0014]

With the above-mentioned configuration, the low boiling point refrigerant does not easily complete its condensation near the condensation completion point of the non-azeotropic mixed refrigerant condenser, and it remains as bubbles forever. In the first means, the fin downstream of the condenser, that is, the fin near the condensation completion point is used as the high-performance fin 3 to accelerate the completion of condensation, so that even if the temperature difference from the cooling air becomes small, A sufficient amount of heat for exchange can be secured, and the condensation completion point can be advanced. As a result, condensation can be completed with a condenser of the same size as in the case of a single refrigerant.

In the second means, the small diameter heat transfer tube 4 is used near the condensation completion point in order to improve the heat transfer coefficient on the refrigerant downstream side of the condenser, that is, near the condensation completion point.
Since the completion of the condensation is promoted, the flow velocity is increased, so that the condensation heat transfer coefficient αs and the liquid heat transfer coefficient αl shown in FIG. 9 are improved, the condensation completion point is accelerated, and the low boiling point component which is likely to remain The bubbles can also quickly become a condensate due to the high flow rate and stirring action generated by the small diameter tube.

[0016]

Embodiments of the present invention will be described below by taking the case where the non-azeotropic mixed refrigerant condenser of the present invention is applied to a room air conditioner or a package air conditioner as an example.

A first embodiment of the present invention will be described with reference to FIG. FIG. 1 is a front view of the non-azeotropic mixed refrigerant condenser of the present embodiment.

As shown in FIG. 1, the gas of the non-azeotropic mixed refrigerant flows into the heat transfer tube 1 provided with the fins 2 from the refrigerant inlet 5, is liquefied and flows out from the refrigerant outlet 6. On the other hand, the cooling air flows almost along the fins 2, that is, perpendicular to the heat transfer tubes 1 (not shown). The superheated gas flowing from the refrigerant inlet 5 is gradually cooled by the cooling air flowing between the fins 2 and reaches the saturation region. In the saturation region, since the refrigerant is a non-azeotropic mixed refrigerant, the condensation proceeds while the temperature gradually decreases as shown by a in FIG. The temperature difference from the cooling air at the time of completion of condensation is ΔTb in the case of a single refrigerant,
In the case of a non-azeotropic mixed refrigerant, it decreases to ΔTa.

Further, as shown in FIG. 8, the latent heat tends to increase as the condensation progresses. In the case of a single refrigerant, the latent heat is constant (indicated by b in FIG. 8) from the condensation start point to the condensation completion point, but in the case of a non-azeotropic mixed refrigerant, the latent heat is indicated by a. Will gradually increase. This is because, as shown in FIG. 6, as the condensation progresses, the molar component ratio of the gas phase changes above AC, that is, above the curve g.

Further, as shown in FIG. 9, the heat transfer coefficient on the refrigerant side may be decreased. As the condensation progresses, the amount of liquid increases and the flow velocity decreases, so the heat transfer coefficient decreases like αs even in the condensation region, but when the condensation is completed and becomes the liquid region, the heat transfer coefficient further decreases to αl. .

Due to these three factors, the non-azeotropic mixed refrigerant is less likely to be condensed than the single refrigerant, and the condensation completion point tends to shift to the downstream side in the refrigerant flow direction. 10. In FIG. 10, the horizontal axis represents the heat transfer tube length of the condenser, and the vertical axis represents the dryness. The non-azeotropic mixed refrigerant (denoted by a in FIG. 10) and the single refrigerant (denoted by b in FIG. 10). Shows the change of. The non-azeotropic mixed refrigerant requires a longer heat transfer tube length until the dryness becomes zero. That is, the required size of the condenser becomes large.

In order to solve this, in the condenser for non-azeotropic mixed refrigerant of this embodiment, as shown in FIG. 1, the normal fins 2 are used on the upstream side of the refrigerant and the high-performance fins 3 are used on the downstream side of the refrigerant. I am using. Although it is conceivable to configure all the fins with the high-performance fins 3, the high-performance fins usually have a drawback that ventilation resistance also increases, and as a result, noise also increases. Further, when this condenser is used as an outdoor heat exchanger of a heat pump, the fins may be frosted during the heating operation in winter. It is preferable to combine high-performance fins because frost easily adheres to the high-performance fins and clogging due to frost easily occurs. From the above, in this embodiment, as shown in FIG. 1, the range in which the high-performance fins 3 are used is limited to the minimum required area on the downstream side of the refrigerant.

A second embodiment of the present invention will be described with reference to FIGS. FIG. 2 is a front view of the non-azeotropic mixed refrigerant condenser of the present embodiment, and FIGS. 3 and 4 are horizontal cross-sectional views of the heat transfer tubes.

In the condenser for non-azeotropic mixed refrigerant of this embodiment, as shown in FIG. 2, the heat transfer tube 1 having a normal diameter is used on the refrigerant inlet side, and the small diameter tube 4 is used only near the refrigerant outlet. It is characterized by having been. Although these heat transfer tubes may be smooth tubes, it is desirable to use tubes with inner grooves as shown in FIGS. 3 and 4. In FIGS. 3 and 4, 1 is an inner grooved tube having a normal diameter, and 4 is a small diameter inner grooved tube having a smaller diameter.

The reason will be described with reference to FIG. In the vicinity of the condensation start point, the gas occupies a large proportion and the flow velocity is high. Therefore, the pressure loss increases when a small-diameter tube is used. Since the flow velocity is also reduced, the condensation heat transfer coefficient αs and the single-phase flow heat transfer coefficient αl in the liquid region can be improved by using the small diameter tube 4. In the region close to the liquid region, the flow velocity is still low even with a small-diameter pipe, so there is no need to worry about the increase in pressure loss.

As described above, by configuring as in the first and second embodiments, the performance of the non-azeotropic mixed refrigerant condenser is improved, and the size of the condenser is almost the same as that of a single refrigerant. The desired subcool can be obtained with the Sano condenser.

FIG. 5 shows a third embodiment of the present invention, in which a pipe array may be provided in two rows to form a condenser, and the same effects as those of the above-described embodiments can be obtained.

[0028]

As described above, according to the present invention, in the condensation process of a non-azeotropic mixed refrigerant, it is possible to improve the condensation performance in the vicinity of the completion point of condensation, and it is possible to obtain a predetermined refrigerant with a volume substantially the same as that of a single refrigerant. You can get subcool.

As a result, the volume of the condenser can be suppressed to the same level as in the case of using a single refrigerant, so that the size of the unit need not be increased, and the efficiency of the entire cycle (also referred to as COP) is reduced. The effect is that you do not have to do.

As a secondary effect, a sufficient subcool (supercooling degree) can be obtained before the expansion valve.
There is an effect that it is possible to obtain the advantage that the refrigerant noise generated in the expansion valve can be reduced.

[0031]

[Brief description of drawings]

FIG. 1 is a front view of a condenser for a non-azeotropic mixed refrigerant which is a first embodiment of the present invention.

FIG. 2 is a front view of a non-azeotropic mixed refrigerant condenser according to a second embodiment of the present invention.

3 is a cross-sectional view of the heat transfer tube 1 shown in FIG.

4 is a cross-sectional view of the small diameter heat transfer tube 4 shown in FIG.

FIG. 5 is a side view of a non-azeotropic mixed refrigerant condenser according to a third embodiment of the present invention.

FIG. 6 is a vapor-liquid equilibrium diagram of a non-azeotropic mixed refrigerant.

FIG. 7 is a diagram showing a refrigerant temperature change in a refrigerant flow direction in a condenser.

FIG. 8 is a diagram showing changes in the latent heat of the refrigerant along the length of the condensation section.

FIG. 9 is a diagram showing changes in the heat transfer coefficient of the refrigerant along the length of the condenser.

FIG. 10 is a diagram showing a change in dryness along a length of a heat transfer tube of a condenser.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Heat transfer tube, 2 ... Fin, 3 ... High performance fin, 4 ... Small diameter heat transfer tube, 5 ... Refrigerant gas inlet, 6 ... Refrigerant liquid outlet, 7 ... Cooling air inlet, 8 ... Cooling air outlet.

Continuation of front page (51) Int.Cl. ⁵ Identification number Internal reference number for FI FI technical display location F25B 1/00 395 A 8919-3L (72) Inventor Hiroaki Matsushima 502 Kitsutachi-cho, Tsuchiura-shi, Ibaraki Hitate Co., Ltd. Machinery Research Laboratory

Claims (2)

[Claims] 1. A condenser of a refrigeration cycle using a non-azeotropic mixed refrigerant in which two or more kinds of refrigerants are mixed, wherein the fins on the upstream side of the refrigerant having a high dryness are more downstream than those on the downstream side of the refrigerant having a lower dryness. A condenser for a non-azeotropic mixed refrigerant, characterized in that the fins are composed of high-performance fins. 2. A condenser of a refrigeration cycle using a non-azeotropic mixed refrigerant in which two or more kinds of refrigerants are mixed, the refrigerant having a lower dryness than the heat transfer tube on the upstream side of the refrigerant having a higher dryness. A condenser for a non-azeotropic mixed refrigerant, characterized in that the heat transfer tube of is composed of a heat transfer tube having a smaller diameter.