JP2006017433A - Air conditioner - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To further improve efficiency of an air conditioner. <P>SOLUTION: The air conditioner 1 is provided with an outdoor side cycle 8 having a compressor 2, a four way valve 3, an intermediate heat exchanger 6 passing a refrigerant and a conveyance medium through an interior, an expansion valve 5, and an outdoor heat exchanger 4, and an indoor side cycle 13 having a conveyance medium drive unit 10 circulating the conveyance medium pressurized into a supercritical zone, the intermediate heat exchanger 6, and at least one indoor heat exchanger 11. Heat exchange is carried out in the intermediate heat exchanger 6 between the refrigerant circulating in the outdoor side cycle 8, and the conveyance medium circulating in the indoor side cycle 13. A non-azeotropic mixed refrigerant is used as the refrigerant, and the intermediate heat exchanger 6 carries out heat exchange during heating operation between the mixed refrigerant, and the conveyance medium changed by sensible heat as counter currents. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to an air conditioner that pressurizes a carrier medium in a supercritical region, encloses the carrier medium in an indoor cycle, and circulates the carrier medium in the indoor cycle to perform an air conditioning operation.

The carrier medium is pressurized to the supercritical region and enclosed in an indoor cycle (also referred to as a “secondary cycle” or “use-side cycle”), and this carrier medium is circulated in the indoor cycle to perform air conditioning operation. As an air conditioner to be performed, for example, a carrier medium is heated or cooled by heat exchange with a refrigerant circulating in an outdoor cycle (also referred to as “primary side cycle” or “heat source side cycle”). (For example, refer to Patent Document 1).
Japanese Patent Laid-Open No. 9-79684 (see, for example, FIG. 7)

By the way, in recent years, in order to deal with environmental problems such as global warming, it is desired to further improve the energy consumption efficiency (COP) of the air conditioner.
However, in the above-described outdoor cycle of the prior art, when the refrigerant outlet temperature of a compressor operated by single-stage compression is constant, it is necessary to operate with the refrigerant evaporation temperature set to one. For this reason, the differential pressure between the inlet pressure and the outlet pressure becomes large regardless of the air conditioning load, and an operation region in which the compressor operation must be performed becomes larger than necessary. In other words, a large amount of compressor drive power is consumed to compress the refrigerant, resulting in a problem that energy consumption efficiency (COP) is lowered.

  The present invention has been made in view of the above circumstances, and aims to further improve the efficiency of an air conditioner.

The present invention employs the following means in order to solve the above problems.
The air conditioner according to claim 1 is a compressor that compresses a refrigerant, a four-way valve that switches a flow path of the refrigerant discharged from the compressor in accordance with an operation mode, and intermediate heat through which the refrigerant and the transport medium pass. An outdoor cycle having an exchanger, an expansion valve that decompresses and expands the refrigerant to form a low-temperature and low-pressure refrigerant, and an outdoor heat exchanger that exchanges heat between the refrigerant that passes through the interior and the outside air, and the supercritical region. A chamber having a conveyance medium driving device that circulates the pressurized conveyance medium, the intermediate heat exchanger, and at least one indoor heat exchanger that exchanges heat between the conveyance medium passing through the inside and the air blown into the room An air conditioner that exchanges heat between the refrigerant circulating in the outdoor cycle and the transport medium circulating in the indoor cycle in the intermediate heat exchanger,
The refrigerant is a non-azeotropic refrigerant, and the intermediate heat exchanger exchanges heat between the refrigerant and the carrier medium that changes sensible heat as a counter flow during heating operation.

  According to such an air conditioner, the refrigerant is a non-azeotropic mixed refrigerant, and in the intermediate heat exchanger, the non-azeotropic mixed refrigerant and the transport medium that changes sensible heat are exchanged in the counter flow during the heating operation. Due to the temperature slip peculiar to the azeotropic refrigerant mixture, the temperature difference from the sensible heat changing carrier medium can be set small and efficient operation becomes possible. In other words, if the slope of the temperature slip is set so that the carrier medium is substantially parallel to the slope of the temperature at which the sensible heat changes, and the temperature difference is set to the minimum, the operating power on the compressor side is suppressed and the efficiency is reduced. Driving is possible.

In the above air conditioner, it is preferable that the compressor performs multi-stage compression, so that the temperature difference from the carrier medium that changes sensible heat can be set to be even smaller over a wider range.
In this case, the multi-stage compression of the compressor may be one that is extracted from a plurality of locations on the way, or may be one in which a plurality of compressors are arranged in series.

In the above air conditioner, the carrier medium is preferably supercritical fluid carbon dioxide (CO ₂ ).

According to the air conditioner of the present invention, in the intermediate heat exchanger during the heating operation, the non-azeotropic mixed refrigerant having a temperature slip and the transport medium changing the sensible heat are exchanged into a counter flow to exchange heat. If the temperature slip is set so that the temperature of the transport medium is almost parallel to the sensible heat temperature gradient and the temperature difference is minimized, the work on the compressor side is optimized and the compressor drive power is consumed. Since the amount can be reduced, the energy consumption efficiency (COP) of the air conditioner can be improved. In addition, since the refrigerant and the transport medium exchange heat with each other, the outlet temperature of the transport medium can be set high. Accordingly, this also increases the energy consumption efficiency (COP) during the heating operation of the air conditioner. Can be improved.
Furthermore, if the compressor performs multi-stage compression, the temperature difference from the carrier medium that changes sensible heat can be set finely over a wide range, so the energy consumption efficiency (COP) of the air conditioner can be further increased. Can be improved.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
FIG. 1 is a schematic configuration diagram showing an embodiment of an air conditioner according to the present invention. An air conditioner 1 shown in FIG. 1 includes an outdoor cycle 8 having a compressor 2, a four-way valve 3, an outdoor heat exchanger 4, an expansion valve 5, an intermediate heat exchanger 6, and a pipe 7 as main elements, and the intermediate cycle. The heat exchanger 6, the conveyance medium drive device 10, the indoor heat exchanger 11, and the indoor side cycle 13 which has the piping 12 as main elements are comprised.

The compressor 2 is a low-temperature / low-pressure mixed refrigerant gas (a mixed refrigerant is a mixture of pure substance refrigerants in a certain composition ratio. For example, a plurality of hydrocarbon refrigerants such as propane, butane, and isobutane (carbon (C ) And hydrogen (H) -containing refrigerant) are sucked and compressed, and sent out as a high-temperature and high-pressure gaseous mixed refrigerant.
The four-way valve 3 is provided on the downstream side of the compressor 2 and switches the flow path of the mixed refrigerant discharged from the compressor 2 (hereinafter abbreviated as “refrigerant”) between the cooling operation and the heating operation. In the heating operation (when the refrigerant is circulated in the direction indicated by the solid line arrow in the figure), a flow path from the compressor 2 to the intermediate heat exchanger 6 is formed, while in the cooling operation (the broken line in the figure) When the refrigerant is circulated in the direction indicated by the arrow), a flow path from the compressor 2 toward the outdoor heat exchanger 4 is formed.
The outdoor heat exchanger 4 functions as a condenser that condenses and liquefies high-temperature and high-pressure gaseous refrigerant during cooling operation and dissipates heat to the outside air, and conversely evaporates low-temperature and low-pressure liquid refrigerant during heating operation and takes heat from the outside air. It functions as an evaporator.
The expansion valve 5 has a function of depressurizing and expanding the refrigerant passing through the inside to form a low-temperature and low-pressure refrigerant.

The intermediate heat exchanger (also referred to as “inter-refrigerant heat exchanger”) 6 allows heat to be exchanged between the refrigerant circulating in the outdoor cycle 8 and the transport medium circulating in the indoor cycle 13. During the heating operation, the transport medium circulating in the indoor cycle 13 is heated by the refrigerant circulating in the outdoor cycle 8, and during the cooling operation, the transport medium circulating in the indoor cycle 13 is cooled by the refrigerant circulating in the outdoor cycle 8. It has come to be. Further, a temperature sensor T is provided in the pipe 7 positioned near the inlet of the intermediate heat exchanger 6, and refrigerant temperature data in the pipe 7 detected by the temperature sensor T is output to the controller 14. It is like that.
The pipe 7 connects the compressor 2, the four-way valve 3, the outdoor heat exchanger 4, the expansion valve 5 and the intermediate heat exchanger 6, and allows the refrigerant to circulate between these components. Thus, an outdoor cycle (also referred to as “primary side cycle” or “heat source side cycle”) 8 is formed.

The carrier medium driving device 10 is a fluid pressure-feeding means such as a pump or a compressor, for example, and sends a supercritical carrier medium (for example, carbon dioxide (CO ₂ )), for example, in the indoor cycle 13. It is something to circulate in.
The indoor heat exchanger 11 performs heat exchange between the transport medium passing through the inside and the air blown into the room, and functions as a so-called evaporator during cooling operation and as a so-called condenser during heating operation. .
The pipe 12 connects the intermediate heat exchanger 6, the indoor heat exchanger 11, and the transport medium driving device 10, and enables the transport medium to circulate between these constituent elements. (Also referred to as “secondary cycle” or “use cycle”) 13 is formed.
The pipes 12 positioned on the upstream side of the indoor heat exchangers 11 are each provided with a flow rate adjusting valve 16 to adjust the amount of the transport medium flowing into the indoor heat exchanger 11 from the intermediate heat exchanger 6. It has become so.

Now, the non-azeotropic refrigerant mixture used in the outdoor cycle 8 described above has a characteristic that is not found in a single refrigerant called temperature slip or temperature gradient. This temperature slip is a phenomenon in which when the non-azeotropic refrigerant mixture undergoes a condensation / evaporation phase change under a constant pressure, the temperature at the start and at the end is different and does not become constant. That is, the mixed refrigerant generally called “non-azeotropic mixed refrigerant” has a temperature difference between the saturated gas temperature and the saturated liquid temperature at a certain pressure. (On the other hand, the mixed refrigerant in which the saturated gas temperature and the saturated liquid temperature are the same is called an “azeotropic mixed refrigerant”.)
A specific example of the non-azeotropic refrigerant mixture is a mixture of propane and isobutane at 50 wt%, and a temperature slip of 5 ° C. occurs. Other specific examples of the non-azeotropic refrigerant mixture include R410A, R404A, and R407C.

Here, the phenomenon of temperature slip described above will be briefly described with reference to the Mollier diagram (ph diagram) shown in FIG.
The low-temperature and low-pressure gaseous non-azeotropic refrigerant mixture compressed by the compressor 2 rises in pressure from point a to point b in the figure to become a high-temperature and high-pressure gas refrigerant, and in the heat exchanger from point b to point c Condensation is caused by a change in the isobaric pressure. At this time, if the temperature of the refrigerant at point b is ti, this refrigerant temperature ti becomes the inlet temperature of the intermediate heat exchanger 6 during the heating operation, and heat is exchanged with the counter-flow carrier medium in the heat exchanger 6. Then, the refrigerant outlet temperature after being condensed and liquefied decreases to to. In the process of such a temperature change (ti → to), a temperature difference is generated between the saturated gas temperature t1 at which condensation starts and the saturated liquid temperature t2 at which condensation is completed, and this temperature difference is called a temperature slip. Therefore, the average saturation temperature t of the non-azeotropic refrigerant mixture whose state changes in the intermediate heat exchanger 6 is obtained by the following equation.
t = (t1 + t2) / 2

  Moreover, the broken line display in a figure has shown the case of a single refrigerant | coolant, and it is necessary to pressurize to the point b 'with the compressor 2 in order to set it as the same saturated gas temperature t1. That is, it can be seen that in order to achieve the same saturated gas temperature t1, it is necessary to pressurize to a point b 'higher than the point b of the mixed refrigerant.

By the way, the above-described transport medium for supercritical fluid has a region where a large amount of heat (enthalpy) change can be obtained even with a slight temperature difference. FIG. 3 is an isobaric diagram showing the characteristics of carbon dioxide as an example of a supercritical fluid. In the isobaric line from the critical pressure limit to around 10 MPa, the temperature rise is not proportional to the increase in enthalpy. This is because the isobaric strain remains in the supercritical region due to the influence of the two-phase region (latent heat region) up to 7.6 MPa.
In such a region, the isobaric line is concave, and the slope of the curve is very small and there is a part close to the horizon (low slope part). By using this part effectively, the temperature difference from the refrigerant is small. Even in exchange, a large change in heat quantity can be obtained. By the way, when the supercritical fluid is carbon dioxide, if the practical minimum temperature level is set to 50 ° C or less, a large amount of heat changes from a small temperature difference using the low slope part that exists in the concave curve part of the isobaric line. Therefore, the outlet temperature of the intermediate heat exchanger 6 at the time of heating operation becomes high, which is effective for improving COP.

FIG. 4 shows the heat exchange (temperature change) between the refrigerant and the transfer medium in the intermediate heat exchanger 6 using the above-described low-inclined portion of the supercritical fluid, and according to the heat exchange distance on the horizontal axis. The changing temperature is shown on the vertical axis. In the figure, the primary side is a non-azeotropic refrigerant mixture, the secondary side is a carrier fluid (supercritical fluid) that changes sensible heat, and the temperature change of a single refrigerant is shown by an imaginary line as a reference example. .
FIG. 4 shows a case where the compressor 2 is single-stage compression, and in the case of heating operation in the counterflow, the primary side flows from the right (inlet) to the left (outlet) in the figure, and the temperature due to heat exchange The change (temperature decrease) and state change occur, and the secondary side flows from the left (inlet) to the right (exit) in the figure to change the temperature (temperature increase). In this case, since the transport medium on the secondary side is a sensible heat change, the temperature is raised by absorbing heat from the primary side having a higher temperature as the heat exchange distance increases.

  The primary non-azeotropic refrigerant mixture serving as a heat source is supplied from the compressor 2 to the intermediate heat exchanger 6 in a high-temperature and high-pressure gas state (gas phase), and then exchanges heat with the secondary-side transport medium. The liquid refrigerant (liquid phase) flows out of the heat exchanger. That is, the mixed refrigerant introduced into the intermediate heat exchanger 6 in the gas state at the inlet temperature ti flows out in the liquid state at the outlet temperature to. At this time, a gas-liquid two-phase non-azeotropic refrigerant mixture having a temperature slip exists in the intermediate heat exchanger 6. The non-azeotropic refrigerant mixture starts condensing at the saturated gas temperature t1 and completes condensing at the saturated liquid temperature t2. The secondary-side transport medium rises in temperature due to the heat of condensation heat generated by the condensation having such a temperature slip.

Therefore, the non-azeotropic refrigerant mixture that condenses in the intermediate heat exchanger 6 has a temperature gradient that is inclined in the same direction as the transport medium on the secondary side, and the two inclinations can be set substantially in parallel. That is, when compared with a single refrigerant having a constant condensation temperature indicated by an imaginary line, a non-azeotropic refrigerant mixture having a temperature slip (temperature gradient) substantially parallel to the temperature change on the secondary side that rises to the right is selected. Thus, heat exchange can be performed with the temperature difference set to a substantially minimum over a wide range in the intermediate heat exchanger 6.
In the illustrated example, the primary side condensation start temperature (saturated gas temperature) is set to t1, and the secondary side outlet temperature is determined according to the setting of the condensation start temperature t1. That is, the outlet temperature on the secondary side is substantially determined by the temperature difference (Δt) between the condensation temperature t1 at the dew point and the supercritical fluid temperature on the endothermic side.

  In counterflow heat exchange employing such a non-azeotropic refrigerant mixture, compared to the case where the single refrigerant is set to the same inlet temperature ti as that of the non-azeotropic refrigerant mixture, in other words, the single refrigerant is not Compared with the case where the saturated gas temperature t1 is set to be the same as that of the azeotropic refrigerant mixture, the pressure compressed by the compressor 2 can be lowered. That is, when the same saturated gas temperature t1 is set, as shown in FIG. 2, the pressure applied by the compressor 2 is lower in the non-azeotropic refrigerant mixture than in the single refrigerant, so that the compressor 2 is driven accordingly. Power can be reduced. Therefore, the energy consumption of the compressor 2 can be reduced and the COP of the air conditioner 1 can be improved.

  Further, when the average saturation temperature t of the non-azeotropic refrigerant mixture is made the same as the saturation temperature t1 of the single refrigerant (see FIG. 2), that is, when the compression pressure by the compressor 2 is made the same, there is a non-temperature slippage. The azeotropic refrigerant mixture has a higher inlet temperature ti and saturated gas temperature t1 than a single refrigerant. For this reason, even if the energy consumption amount of the compressor 2 is the same, the intermediate heat exchanger 6 can heat the transport medium with a high-temperature refrigerant, and thus can be heated to a higher temperature. That is, since the transport medium can be heated to a higher temperature with the same energy consumption as that of the single refrigerant, the COP of the air conditioner 1 can be improved.

In the above-described embodiment, the compressor 2 of the outdoor cycle 8 is single-stage compression. However, as in the other embodiment shown in FIG. Good. In this configuration example, a plurality (three in the illustrated example) of compressors C1, C2, and C3 are arranged in series. Note that the flow direction of the refrigerant and the carrier medium in this case indicates the state during the heating operation, and in the intermediate heat exchanger 6, the opposite flows are opposite to each other.
In the illustrated compressor 2 ′, three compressors C 1, C 2 and C 3 are arranged in series from the high-pressure stage side, but a low-pressure stage compressor that introduces low-temperature and low-pressure gas refrigerant from the outdoor heat exchanger 4. The discharge side pipe 7a of C3 is connected to the intermediate pressure stage compressor C2, and the discharge side pipe 7b of the intermediate pressure stage compressor C2 is connected to the high pressure stage compressor C1. A pipe 7 connected to the discharge side of the high-pressure compressor C1 is connected to a high temperature region 6a in the intermediate heat exchanger 6 as a discharge side pipe of the compressor 2.

Further, in the above-described multistage compression compressor 2, the bypass pipe P1 branched from the discharge side pipe 7a of the intermediate pressure stage compressor C2 and the bypass pipe P2 branched from the discharge side pipe 7b of the low pressure stage compressor C3 are intermediate. It is introduced into the heat exchanger 6. The bypass pipe P1 branched from the intermediate pressure stage compressor C2 is connected to the intermediate temperature region 6b in the intermediate heat exchanger 6 on the upstream side (outside of the intermediate heat exchanger 6) from the high pressure stage compressor C1. A pipe 7c for flowing the refrigerant after being introduced into the intermediate heat exchanger 6 and exchanging heat with the transport medium is connected via an electromagnetic on-off valve 9a.
Similarly, in the middle of the upstream side where the bypass pipe P2 branched from the low pressure stage compressor C3 is connected to the low temperature region 6c in the intermediate heat exchanger 6, the intermediate pressure stage compressor C2 enters the intermediate heat exchanger 6. A pipe 7d for flowing the refrigerant after being introduced and exchanging heat with the transport medium is connected via an electromagnetic on-off valve 9b.
In addition, an electromagnetic on-off valve 9c is disposed in the pipe 7 that connects the intermediate heat exchanger 6 and the outdoor heat exchanger 4.

The multistage compression compressor 2 ′ configured as described above selects the discharge pressure of the high-temperature and high-pressure gas refrigerant supplied to the intermediate heat exchanger 6 by performing the operation described below. Is possible.
First, the case where one low-pressure compressor C3 is operated and the refrigerant discharged at the lowest pressure is supplied to the intermediate heat exchanger 6 will be described. In such an operation, both the electromagnetic on-off valves 9a and 9b are closed, and only the remaining electromagnetic on-off valve 9c is opened. As a result, the low-temperature and low-pressure gas refrigerant introduced from the outdoor heat exchanger 4 is increased in pressure by the low-pressure stage compressor C3 to become a relatively low-pressure gas refrigerant (mass flow rate G3), and is bypassed from the low-pressure stage compressor C3. It is supplied to the low temperature region 6c of the intermediate heat exchanger 6 through P2.

  Next, the case where the two compressors C2 and C3 are operated and the refrigerant discharged at the intermediate pressure is supplied to the intermediate heat exchanger 6 will be described. In such operation, the electromagnetic on-off valve 9a is closed and the other electromagnetic on-off valves 9b and 9c are opened. As a result, the low-temperature and low-pressure gas refrigerant introduced from the outdoor heat exchanger 4 is pressurized by the low-pressure stage compressor C3, and then a part of the gas refrigerant (mass flow rate G2) passes through the intermediate-pressure stage compressor C2 to be 2 Boosted in stages. The gas refrigerant (mass flow rate G2) having the intermediate pressure in this way is supplied from the compressor C2 through the bypass pipe P1 to the intermediate temperature region 6b of the intermediate heat exchanger 6, and further directly supplied from the low-pressure compressor C3. Gas refrigerant (mass flow rate G3) joins, and both gas refrigerants (mass flow rate G2 + G3) pass through the low temperature region 6c and exchange heat with the transport medium. In this case, since the refrigerant and the carrier medium are counterflowed, the carrier medium is first heated from the inlet temperature by exchanging heat with the refrigerant (mass flow rate G2 + G3) having a relatively low temperature in the low temperature region 6c. The temperature is raised by exchanging heat between the higher temperature gas refrigerant (mass flow rate G2) and the intermediate temperature region 6b. For this reason, when heat is exchanged in the intermediate temperature region 6b, the transfer medium is a one-step temperature increase from the inlet temperature of the intermediate heat exchanger 6, so that efficient heat exchange is possible.

  Finally, the case where all the three compressors C1, C2, and C3 are operated and the refrigerant having the highest discharge pressure is supplied to the intermediate heat exchanger 6 will be described. In such an operation, the electromagnetic on-off valves 9a, 9b, 9c are all opened, and the low-temperature and low-pressure gas refrigerant introduced from the outdoor heat exchanger 4 is partially pressurized after being pressurized by the low-pressure stage compressor C3. Passes through the intermediate-pressure stage compressor C2 and the high-pressure stage compressor C3, and the pressure is sequentially increased. In this way, the high-temperature and high-pressure gas refrigerant (mass flow rate G1) is supplied from the compressor C1 to the intermediate heat exchanger 6, passes through all of the high-temperature region 6a, the intermediate temperature region 6b, and the low-temperature region 6c in this order, and exchanges heat with the transport medium. For the intermediate temperature region 6b, the gas refrigerant (mass flow rate G2) directly supplied from the intermediate pressure stage compressor C2 is added, and the merged gas refrigerant (mass flow rate G1 + G2) passes to exchange heat with the transport medium. . Further, in the low temperature region 6c, the gas refrigerant (mass flow rate G3) directly supplied from the low-pressure compressor C3 is added, and the merged gas refrigerant (mass flow rate G1 + G2 + G3) passes to exchange heat with the transport medium.

  At this time, since the refrigerant and the conveyance medium are counterflowed, the conveyance medium is first heated from the inlet temperature by exchanging heat with the refrigerant (mass flow rate G1 + G2 + G3) having a relatively low temperature in the low temperature region 6c. The temperature is increased by heat exchange with the gas refrigerant (mass flow rate G1 + G2) and the gas refrigerant (mass flow rate G1) in the intermediate temperature region 6b and the high temperature region 6a having successively higher temperatures. For this reason, when heat exchange is performed in the high temperature region 6a having the highest refrigerant temperature, the transfer medium is a relatively high-temperature transport medium that is two-stage temperature rise from the inlet temperature of the intermediate heat exchanger 6, and thus efficient heat exchange. Becomes possible, and the outlet temperature of the transport medium can be taken out at a higher temperature.

If a non-azeotropic refrigerant mixture is employed for the above-described multistage (three-stage) compression, as shown in FIG. 6, when the secondary side outlet temperature is set to the same value as the above-described single-stage compression, The condensation temperature can be set to a plurality of (three stages) temperature levels in accordance with the sensible heat change. In the illustrated example, the condensation temperature of the high-pressure compressor C1 that supplies the gas refrigerant with the mass flow rate G1 is set to substantially the same value as that of the single-stage compression in FIG. 4, and the intermediate-pressure compressor C2 and the low-pressure compressor C3 It is set so that the condensing temperature becomes lower step by step. The condensation temperature of the refrigerant has a correlation with the pressure, and the higher the pressure, the higher the temperature.
And, by the temperature slip of the non-azeotropic refrigerant mixture, the primary side slope can be matched to the secondary side temperature rise at each compression stage or for each compression stage, so compared with single stage compression, A setting closer to the secondary side temperature rise over a wide heat exchange region is possible. In other words, since it is possible to perform fine operation in which the temperature difference is set to be small so as to be substantially parallel to the temperature change of the transport medium, the driving force consumed for the operation of the compressor 2 ′ can be reduced and efficient operation can be achieved. It becomes possible.

  As described above, in the air conditioner 1 of the present invention, the work of the compressor 2 can be optimized by effectively utilizing the temperature slip of the non-azeotropic refrigerant mixture. Consumption can be reduced and COP can be improved. Further, since the non-azeotropic refrigerant mixture and the transport medium exchange heat with each other, the outlet temperature of the transport medium can be set high, which also improves the COP during the heating operation of the air conditioner 1. Can be made.

In the above-described multi-stage compression, the compressors are arranged in series and the compressor 2 'performs the three-stage compression. However, the multi-stage compression may be performed in two stages or four or more stages. For multi-stage compression, in addition to arranging a plurality of compressors in series, for example, a plurality of stages with different discharge pressures and condensation temperatures may be formed by extracting air at a plurality of locations from the middle of the scroll compressor. Good.
The present invention is not limited to the above-described embodiment, and can be changed by appropriately selecting the primary-side refrigerant and the secondary-side transport medium, as long as they do not depart from the gist of the present invention. It can be changed as appropriate.

It is a schematic block diagram which shows one Embodiment of the air conditioning apparatus by this invention. It is a Mollier diagram (ph diagram) for explaining a temperature slip phenomenon of a non-azeotropic refrigerant mixture. It is an isobaric diagram which showed the characteristic of the carbon dioxide as an example of a supercritical fluid. It is a figure which shows the case of single stage compression about the heat exchange (temperature change) of the primary side (refrigerant) and secondary side (conveyance medium) in an intermediate heat exchanger. It is a figure which shows other embodiment of this invention, and is a schematic block diagram which shows the case where the compressor of an outdoor cycle is multistage compression. It is a figure which shows the case of three-stage compression about the heat exchange (temperature change) of the primary side (refrigerant) and secondary side (conveyance medium) in an intermediate heat exchanger.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Air conditioning apparatus 2 Compressor 4 Outdoor heat exchanger 5 Expansion valve 6 Intermediate heat exchanger 8 Outdoor cycle 10 Transport medium drive device 11 Indoor heat exchanger 13 Indoor cycle

Claims (3)

A compressor that compresses the refrigerant, a four-way valve that switches the flow path of the refrigerant discharged from the compressor according to the operation mode, an intermediate heat exchanger through which the refrigerant and the transport medium pass, and a low temperature by decompressing and expanding the refrigerant An outdoor valve having an expansion valve for making a low-pressure refrigerant, and an outdoor heat exchanger for exchanging heat between the refrigerant passing through the inside and the outside air;
At least one indoor heat that exchanges heat between the transport medium driving device that circulates the transport medium pressurized in the supercritical region, the intermediate heat exchanger, and the transport medium that passes through the interior and the air that is blown into the room An indoor cycle having an exchanger,
An air conditioner for exchanging heat between the refrigerant circulating in the outdoor cycle and the carrier medium circulating in the indoor cycle in the intermediate heat exchanger,
An air-conditioning apparatus, wherein the refrigerant is a non-azeotropic refrigerant mixture, and the intermediate heat exchanger exchanges heat between the refrigerant and the carrier medium that changes sensible heat as a counterflow during heating operation.   The air conditioner according to claim 1, wherein the compressor performs multi-stage compression. The air conditioning apparatus according to claim 1, wherein the carrier medium is carbon dioxide (CO ₂ ) as a supercritical fluid.

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