Background Art FIG. 10 is a flow diagram of a conventional air- conditioning system. As shown in FIG. 10, there has heretofore been available a dehumidifying air-conditioning apparatus having a compressor 201 for compressing a refrigerant, a condenser 202 for condensing the compressed refrigerant with outside air OA, an evaporator 204 for depressurizing the condensed refrigerant with an expansion valve 203 and evaporating the refrigerant to cool process air from an air-conditioned space 100 to a temperature lower than its dew point, and a reheater 205 for reheating the process air, which has been cooled to a temperature lower than its dew point, at the downstream side of the condenser 202 with the refrigerant upstream of the expansion valve 203.

The refrigerant is condensed in the condenser 202 and the reheater 205. With the illustrated dehumidifying air- conditioning apparatus, a heat pump HP is constituted by the compressor 201, the condenser 202, the reheater 205, the expansion valve 203, and the evaporator 204. The heat pump HP pumps heat from the process air which flows through the evaporator 204 into the outside air OA which flows through the condenser 202.

FIG. 11 is a Mollier diagram in the case where HFC134a is used as the refrigerant in the conventional dehumidifying air-conditioning apparatus. In FIG. 11, a point a represents a state of the refrigerant evaporated by the evaporator 204, and the refrigerant is in the form of a saturated vapor. The refrigerant has a pressure of 0.34 MPa, a temperature of 5°C, and an enthalpy of 400.9 kJ/kg. A point b represents a state of the vapor drawn and compressed by the compressor 201, i. e., a state at the outlet port of the compressor 201. In the point b, the refrigerant is in the form of a superheated vapor.

The refrigerant vapor is cooled in the condenser 202 and reaches a state represented by a point c in the Mollier diagram. In the point c, the refrigerant is in the form of a saturated vapor and has a pressure of 0.94 MPa and a temperature of 38°C. Under this pressure, the refrigerant is cooled and condensed to reach a state represented by a point d. In the point d, the refrigerant is in the form of a saturated liquid and has the same pressure and temperature as those in the point c. The saturated liquid has an enthalpy of 250.5 kJ/kg.

The refrigerant liquid is depressurized by the expansion valve 203 to a saturation pressure of 0.34 MPa at a temperature of 5°C and reaches a state represented by the point e. The refrigerant at the point e is delivered as a mixture of the refrigerant liquid and the vapor at a temperature of 5'C to the evaporator 204, in which the mixture removes heat from process air and is evaporated to reach a state of the saturated vapor, which is represented by the point a in the Mollier diagram. The saturated vapor is drawn into the compressor 201 again, and the above cycle is repeated.

FIG. 12 is a psychrometric chart showing an air- conditioning cycle in the conventional dehumidifying

air-conditioning apparatus. In FIG. 12, the alphabetical letters K, L, M correspond to the encircled letters in FIG. 10. As shown in FIG. 12, in the conventional dehumidifying air-conditioning apparatus, air (in a state K) from the air-conditioned space 100 is cooled to a temperature lower than its dew point to lower the dry bulb temperature thereof and lower the absolute humidity thereof, and reaches a state L. The state L is on a saturation curve in the psychrometric chart. The air in the state L is reheated by the reheater 205 to increase the dry bulb temperature thereof and keep the absolute humidity thereof constant, and reaches a state M. Then, the air is supplied to the air-conditioned space 100. The state M is lower in both of absolute humidity and dry bulb temperature than the state K.

With the conventional dehumidifying air-conditioning apparatus described above, since it is necessary to considerably cool the air to its dew point, about half of the cooling effect of the evaporator in the heat pump is consumed to remove a sensible heat load from the air, so that the moisture removal (the dehumidifying performance) per electric power consumption is low. If a single-stage compressor is used as the compressor in the heat pump, then it produces a one-stage compression-type refrigerating cycle, resulting in a low coefficient of performance (COP) and a large amount of electric power consumed per amount of moisture removal Disclosure of Invention The present invention has been made in view of the above drawbacks. It is therefore an object of the present invention to provide a heat pump with a high coefficient of performance (COP) and a dehumidifying air-conditioning apparatus which consumes a small amount of energy per amount

of moisture removal.

In order to attain the above obj ect, according to a first aspect of the present invention, there is provided a heat pump comprising: a pressurizer for raising a pressure of a refrigerant; a condenser for condensing the refrigerant to heat a high-temperature heat source fluid; an evaporator for evaporating the refrigerant to cool a low-temperature heat source fluid; a refrigerant path branched into a plurality of branched refrigerant paths between the condenser and the evaporator; a first heat exchanging portion disposed in the branched refrigerant path between the condenser and the evaporator for evaporating the refrigerant under an intermediate pressure between the condensing pressure of the condenser and the evaporating pressure of the evaporator to cool the low-temperature heat source fluid by evaporation of the refrigerant under the intermediate pressure; a second heat exchanging portion disposed in the branched refrigerant path between the condenser and the evaporator for condensing the refrigerant under an intermediate pressure between the condensing pressure of the condenser and the evaporating pressure of the evaporator to heat the low-temperature heat source fluid by condensation of the refrigerant under the intermediate pressure; a low-temperature heat source fluid path connecting the first heat exchanging portion, the evaporator, the second heat exchanging portion in the order named.

According to a second aspect of the present invention, there is provided a dehumidifying air-conditioning apparatus comprising: a pressurizer for raising a pressure of a refrigerant; a condenser for condensing the refrigerant to heat a high-temperature heat source fluid; an evaporator for evaporating the refrigerant to cool process air to a temperature lower than its dew point; a refrigerant path

branched into a plurality of branched refrigerant paths between the condenser and the evaporator; a first heat exchanging portion disposed in the branched refrigerant path between the condenser and the evaporator for evaporating the refrigerant under an intermediate pressure between the condensing pressure of the condenser and the evaporating pressure of the evaporator to cool the process air by evaporation of the refrigerant under the intermediate pressure; a second heat exchanging portion disposed in the branched refrigerant path between the condenser and the evaporator for condensing the refrigerant under an intermediate pressure between the condensing pressure of the condenser and the evaporating pressure of the evaporator to heat the process air by condensation of the refrigerant under the intermediate pressure; a process air path connecting the first heat exchanging portion, the evaporator, the second heat exchanging portion in the order named.

With the above arrangement, the low-temperature heat source fluid can be precooled in the first heat exchanging portion prior to cooling in the evaporator. The low- temperature heat source fluid can be heated in the second heat exchanging portion after cooling in the evaporator with use of the heat in precooling. When process air is used as the low-temperature heat source and is cooled to a temperature lower than its dew point by the evaporator, it is possible to provide a dehumidifying air-conditioning apparatus which consumes a small amount of energy per amount of moisture removal.

Further, with the branched refrigerant paths, the operative temperature of the refrigerant can gradually be changed to achieve a high efficiency of heat exchange. When the high-temperature fluid is cooled, i. e., the heat exchanger is used for cooling the high-temperature fluid,

the efficiency of heat exchange is defined by Q = (TP1-TP2)/(TP1-TC1) where the temperature of the high-temperature fluid at the inlet of the heat exchanger is represented by TP1, the temperature thereof at the outlet of the heat exchanger by TP2, the temperature of the low-temperature fluid at the inlet of the heat exchanger is represented by TC1, and the temperature thereof at the outlet of the heat exchanger by TC2. When the low-temperature fluid is to be heated, i. e., when the heat exchanger is used to heat the low-temperature fluid, the efficiency ¢ of heat exchange is defined by Q (TC2-TC1)/ (TP1-TC1) According to a preferred aspect of the present invention, the branched refrigerant paths extend through the interior of the evaporator in parallel, respectively, and are joined to each other at the downstream side of the evaporator. In this case, there may be provided an ejector on a branched refrigerant path for a refrigerant which exchanges heat with process air having a high temperature, for pressurizing a refrigerant which exchanges heat with process air having a low temperature by a refrigerant which has passed through the branched refrigerant path.

With the above arrangement, since the operative temperature of the evaporator is increased to improve the theoretical cooling effect, the theoretical work of compression is reduced to achieve a high efficiency. Further, the specific volume of the refrigerant is reduced to increase the flow rate of the refrigerant drawn by the pressurizer.

Therefore, an amount of moisture removal is increased according to the improved cooling effect, and hence a high efficiency can be achieved.

Brief Description of Drawings FIG. 1 is a schematic view showing a whole arrangement of an air-conditioning system according to a first embodiment of the present invention; FIG. 2 is a flow diagram of a dehumidifying air- conditioning apparatus according to the first embodiment of the present invention; FIG. 3 is an enlarged view showing branched refrigerant paths in a heat exchanger of the dehumidifying air- conditioning apparatus shown in FIG. 2; FIG. 4A is a perspective view showing a heat exchanger and an evaporator in the case where a refrigerant path is not branched, as viewed from a front side; FIG. 4B is a perspective view showing a heat exchanger and an evaporator in the case where a refrigerant path is not branched, as viewed from a rear side; FIG. 5 is a Mollier diagram of a heat pump included in the dehumidifying air-conditioning apparatus shown in FIG.

2; FIG. 6 is a psychrometric chart showing an air- conditioning cycle in the dehumidifying air-conditioning apparatus shown in FIG. 2 ; FIG. 7 is a graph showing the relationship between the number of branched refrigerant paths and the temperature efficiency in a dehumidifying air-conditioning apparatus according to the present. invention; FIG. 8 is a flow diagram of a dehumidifying air- conditioning apparatus according to a second embodiment of the present invention; FIG. 9 is a Mollier diagram of a heat pump included in the dehumidifying air-conditioning apparatus shown in FIG.

8; FIG. 10 is a flow diagram of a conventional

dehumidifying air-conditioning apparatus; FIG. 11 is a Mollier diagram of a heat pump included in the conventional dehumidifying air-conditioning apparatus; and FIG. 12 is a psychrometric chart showing an air- conditioning cycle in the conventional dehumidifying air-conditioning apparatus.

Best Mode for Carrying Out the Invention A dehumidifying air-conditioning apparatus according to a first embodiment of the present invention will be described below with reference to FIGS. 1 through 6. FIG.

1 is a schematic view showing a whole arrangement of an air-conditioning system according to the first embodiment of the present invention, and FIG. 2 is a flow diagram of a dehumidifying air-conditioning apparatus according to the first embodiment of the present invention. The dehumidifying air-conditioning apparatus in the first embodiment serves to cool process air to a temperature lower than its dew point for dehumidifying the air. The dehumidifying air-conditioning apparatus includes a heat pump HP1 therein. The dehumidifying air-conditioning apparatus lowers the humidity of the process air to maintain a comfortable environment in an air-conditioned space 100 supplied with the process air.

As shown in FIG. 1, the dehumidifying air-conditioning apparatus mainly comprises an indoor unit 10 and an outdoor unit 20 installed outside of the air-conditioned space 100 (outdoor). The indoor unit 10 in the dehumidifying air- conditioning apparatus comprises a refrigerant evaporator 1 for evaporating a refrigerant, a heat exchanger 2 for exchanging heat between the refrigerant and the process air, and an air blower 3 for circulating the process air. The heat

exchanger 2 performs a heat exchange between process air flowing into the evaporator 1 and process air flowing out of the evaporator 1, indirectly with the refrigerant. The heat exchanger 2 has a first heat exchanging portion 21 for evaporating the refrigerant to cool the process air, and a second heat exchanging portion 22 for condensing the refrigerant to heat the process air. The outdoor unit 20 in the dehumidifying air-conditioning apparatus comprises a pressurizer (compressor) 4 for raising a pressure of the refrigerant, a refrigerant condenser 5 for cooling and condensing the refrigerant, and an air blower 6 for circulating the cooling air.

Process air paths, which are paths for circulating process air, include a path 30 connecting the air-conditioned space 100 and the first heat exchanging portion 21 in the heat exchanger 2, a path 31 connecting the first heat exchanging portion 21 and the evaporator 1, a path 32 connecting the evaporator 1 and the second heat exchanging portion 22 in the heat exchanger 2, a path 33 connecting the second heat exchanging portion 22 and the air blower 3, and a path 34 connecting the air blower 3 and the air-conditioned space 100. Thus, the first heat exchanging portion 21 in the heat exchanger 2, the evaporator 1, and the second heat exchanging portion 22 in the heat exchanger 2 are connected in the order named by the process air paths.

Refrigerant paths include a path 40 connecting the evaporator 1 and the compressor 4, a path 41 connecting the compressor 4 and the condenser 5, and a path connecting the condenser 5 and the evaporator 1. The path connecting the condenser 5 and the evaporator 1 is branched into a plurality of branched refrigerant paths at the downstream side of the condenser 5. In FIG. 2, three branched refrigerant paths 42, 43,44 are formed at the downstream side of the condenser

5. The branched refrigerant paths 42,43,44 are joined to one path 45 at the upstream side of the evaporator 1.

Outside air OA is introduced as cooling air through the path 46 into the condenser 5. The outside air OA removes heat from the refrigerant which is condensed, and the heated outside air OA is drawn through the path 47 into the air blower 6, from which the air is discharged through the path 48 as exhaust air EX.

The branched refrigerant paths 42, 43,44 penetrate the first heat exchanging portion 21 and the second heat exchanging portion 22 in the heat exchanger 2, respectively.

An evaporating section 51 for evaporating the refrigerant to cool the process air which flows through the first heat exchanging portion 21 is provided in the first heat exchanging portion 21 of the heat exchanger 2. A condensing section 52 for condensing the refrigerant to heat (reheat) the process air which flows through the second heat exchanging portion 22 is provided in the second heat exchanging portion 22 of the heat exchanger 2. Restrictions 11,12,13 are disposed on the respective branched refrigerant paths 42,43,44 at the upstream side of the first heat exchanging portion 21.

Restrictions 14,15,16 are disposed on the respective branched refrigerant paths 42,43,44 at the downstream side of the second heat exchanging portion 22. The restrictions 11-16 may comprise orifices, capillary tubes, expansion valves, or the like.

FIG. 3 is an enlarged view showing the branched refrigerant paths 42,43,44 in the heat exchanger 2 of the dehumidifying air-conditioning apparatus shown in FIG. 2.

The refrigerant paths including the evaporating section 51 and the condensing section 52 penetrate the first heat exchanging portion 21 and the second heat exchanging portion 22 in the heat exchanger 2, alternately and repeatedly.

Specifically, as shown in FIG. 3, the refrigerant path 42 has an evaporating section 61a, a condensing section 62a, a condensing section 62b, an evaporating section 61b, an evaporating section 61c, and a condensing section 62c. The refrigerant path 43 has an evaporating section 63a, a condensing section 64a, a condensing section 64b, an evaporating section 63b, an evaporating section 63c, and a condensing section 64c. The refrigerant path 44 has an evaporating section 65a, a condensing section 66a, a condensing section 66b, an evaporating section 65b, an evaporating section 65c, and a condensing section 66c.

The heat exchanger 2 has the first heat exchanging portion 21 for allowing the process air before flowing through the evaporator 1 to pass therethrough, and the second heat exchanging portion 22 for allowing the process air after flowing through the evaporator 1 to pass therethrough. The first heat exchanging portion 21 and the second heat exchanging portion 22 form respective separate spaces, each in the form of a rectangular parallelepiped. The evaporator 1 is disposed between the first heat exchanging portion 21 and the second heat exchanging portion 22. FIGS. 4A and 4B show the arrangement of a heat exchanger and an evaporator in the case where a refrigerant path is not branched, for reference. FIG. 4A is a perspective view as viewed from the front side, and FIG. 4B is a perspective view as viewed from the rear side.

The first heat exchanging portion 21 and the second heat exchanging portion 22 have a plurality of substantially parallel heat exchange tubes as refrigerant passages in each of a plurality of planes which lie perpendicularly to the flow of the process air. Tubes 67 are provided across the <BR> <BR> <BR> <BR> evaporator 1 between the corresponding sections, for example, the evaporating section 61a and the condensing section 62a,

the evaporating section 61b and the condensing section 62b, and the evaporating section 61c and the condensing section 62c (see FIG. 4B). Thus, the corresponding evaporating and condensing sections are connected to each other. The ends of the evaporating sections 61b, 61c, the ends of the evaporating sections 63b, 63c, and the ends of the evaporating sections 65b, 65c are connected to each other by a U tube 68. Similarly, the ends of the condensing sections 62a, 62b, the ends of the condensing sections 64a, 64b, and the ends of the condensing sections 66a, 66b are connected to each other by a U tube 69 (see FIG. 4A).

With the above arrangement, for example, in the refrigerant path 42, the refrigerant flowing in one direction from the evaporating section 61a to the evaporating section 62a is introduced into the condensing section 62b by the U tube 69. The refrigerant introduced into the condensing section 62b then flows into the evaporating section 61b, from which the refrigerant flows into the evaporating section 61c via the U tube 68 and further flows into the condensing section 62c. In this manner, the refrigerant passages are provided as a group of meandering thin pipes. A group of meandering thin pipes pass through the first heat exchanging portion 21 and the second heat exchanging portion 22, and are held in alternate contact with the process air which has a higher temperature and the process air which has a lower temperature.

As shown in FIGS. 1 and 2, a drain pan 7 is provided in the indoor unit 10 of the dehumidifying air-conditioning apparatus. The drain pan 7 is preferably located below not only the evaporator 1, but also the heat exchanger 2.

Particularly, the drain pan 7 is preferably disposed below the first heat exchanging portion 21 because the process air is mainly precooled in the first heat exchanging portion 21 and some moisture may possibly be condensed in the first heat

exchanging portion 21.

The flow of the refrigerant in the devices will be described below with reference to FIGS. 2 and 3.

A refrigerant vapor pressurized by the compressor 4 is introduced into the condenser 5 via the refrigerant pipe 41 connected to the discharge port of the compressor 4. The refrigerant vapor compressed by the compressor 4 is cooled and condensed by the outside air OA as cooling air. The refrigerant liquid flowing out of the condenser 5 is branched into the branched refrigerant paths 42,43,44. The refrigerants similarly flow through the respective refrigerant paths 42,43,44. Therefore, the refrigerant flowing through the refrigerant path 42 will mainly be described below, and the refrigerants flowing through the other refrigerant paths 43,44 will not be described in detail below.

The refrigerant flowing through the refrigerant path 42 is depressurized by the restriction 11 and expanded so as to be partly evaporated (flashed). The refrigerant which is a mixture of the liquid and the vapor reaches the evaporating section 61a, where the refrigerant liquid flows so as to wet the inner wall surface of the tube in the evaporating section 61a. The refrigerant flows into the evaporating section 61a in the liquid phase. The refrigerant may be a refrigerant liquid which has been partly evaporated to slightly contain a vapor phase. While the refrigerant liquid is flowing through the evaporating section 61a, it is evaporated to cool (precool) the process air before flowing into the evaporator 1. The refrigerant itself is heated while increasing the vapor phase thereof.

As described above, the evaporating section 61a and the condensing section 62a are constructed as a continuous tube.

Specifically, since the evaporating section 61a and the

condensing section 62a are provided as an integral passage, the refrigerant vapor evaporated in the evaporating section 61a (and the refrigerant liquid which has not been evaporated) flows into the condensing section 62a, and heats (reheats) the process air, which has been cooled and dehumidified in the evaporator 1 and has a temperature lower than the process air in the evaporating section 61a. At this time, heat is removed from the evaporated refrigerant vapor itself, and while the evaporated refrigerant vapor in the vapor phase is condensed, the refrigerant flows into the next condensing section 62b. While the refrigerant is flowing through the condensing section 62b, heat is further removed from the refrigerant by the process air having a lower temperature, and the refrigerant in the vapor phase is further condensed.

The condensed refrigerant liquid flows into the next evaporating section 61b and the subsequent evaporating section 61c to cool (precool) the process air before flowing into the evaporator 1 in the same manner as described above.

Thereafter, the refrigerant vapor flows into the condensing section 62c to heat (reheat) the process air. In this manner, the refrigerant flows through the branched refrigerant path while changing in phase between the vapor phase and the liquid phase. Thus, heat is exchanged between the process air before being cooled by the evaporator 1 and the process air which has been cooled by the evaporator 1 to lower its absolute humidity.

The refrigerant liquid condensed in the condensing section 62c is depressurized and expanded by the restriction 14 provided at the downstream side of the second heat exchanging portion 22, for thereby lowering its pressure.

Then, the refrigerant liquid is joined to the refrigerants which have flowed through the other branched refrigerant liquid paths 43,44. The joined refrigerant liquid enters

the evaporator 1 to be evaporated to cool the process air with heat of evaporation. The refrigerant which has been evaporated into a vapor in the evaporator 1 is introduced into the suction side of the compressor 4 through the path 40, and thus the above cycle is repeated.

Next, operation of the heat pump HP1 included in the dehumidifying air-conditioning apparatus according to the first embodiment of the present invention will be described below with reference to FIG. 5. FIG. 5 is a Mollier diagram of the heat pump HP1 included in the dehumidifying air- conditioning apparatus shown in FIG. 2. The diagram shown in FIG. 5 is a Mollier diagram in the case where HFC134a is used as the refrigerant. In the Mollier diagram, the horizontal axis represents the enthalpy, and the vertical axis represents the pressure. In addition to the above refrigerant, HFC407C and HFC410A are suitable refrigerants for the heat pump and the dehumidifying air-conditioning apparatus according to the present invention. These refrigerants have an operating pressure region shifted toward a higher pressure side than HFC134a.

In FIG. 5, a point a represents a state of the refrigerant which has been evaporated by the evaporator 1 shown in FIG. 2, and the refrigerant is in the form of a <BR> <BR> <BR> <BR> saturated vapor. The refrigerant has a pressure of 0. 234 MPa, a temperature of 5°C, and an enthalpy of 395.1 kJ/kg. Apoint b represents a state of the vapor drawn and compressed by the compressor 4, i. e., a state at the outlet port of the compressor 4. In the point b, the refrigerant has a pressure of 0.706 MPa and is in the form of a superheated vapor.

The refrigerant vapor at the point b is cooled in the condenser 5 and reaches a state represented by a point c in the Mollier diagram. In the point c, the refrigerant is in the form of a saturated vapor and has a pressure of 0.706

MPa and a temperature of 38°C. Under this pressure, the refrigerant is cooled and condensed to reach a state represented by a point d. In the point d, the refrigerant is in the form of a saturated liquid and has the same pressure and temperature as those in the point c. The saturated liquid has an enthalpy of 237.4 kJ/kg.

The refrigerant liquid is branched into the branched refrigerant liquid paths 42,43,44, and the branched refrigerant liquids flow into the heat exchanger 2. First, the refrigerant flowing through the refrigerant path 43 will be described below. The refrigerant liquid is depressurized by the restriction 12 and flows into the evaporating section 63a in the first heat exchanging portion 21. This state is indicated at a point e on the Mollier diagram. The refrigerant liquid is a mixture of the liquid and the vapor because a part of the liquid is evaporated. The pressure of the refrigerant liquid is an intermediate pressure between the condensing pressure in the condenser 5 and the evaporating pressure in the evaporator 1, i. e., is of an intermediate value between 0.234 MPa and 0.706 MPa in the present embodiment.

In the evaporating section 63a, the refrigerant liquid is evaporated under the intermediate pressure, and reaches a state represented by a point fl, which is located intermediately between the saturated liquid curve and the saturated vapor curve, under the intermediate pressure. In the point fl, while a part of the liquid is evaporated, the refrigerant liquid remains in a considerable amount. The refrigerant in the state represented by the point fl flows into the condensing sections 64a, 64b. In the condensing sections 64a, 64b, heat is removed from the refrigerant by the process air which has a low temperature and flows through the second heat exchanging portion 22, and the refrigerant

reaches a state represented by a point gl.

The refrigerant in the state represented by the point gl flows into the evaporating sections 63b, 63c, where heat is removed from the refrigerant. The refrigerant increases its liquid phase and reaches a state represented by a point f2. Then, the refrigerant flows into the condensing section 64c, where the refrigerant increases its liquid phase and reaches a state represented by a point g2. On the Mollier diagram, the point g2 is on the saturated liquid curve. In this point, the refrigerant has a temperature of 18°C and an enthalpy of 209.5 kJ/kg.

The refrigerant liquid at the point g2 is depressurized to 0.234 MPa, which is a saturated pressure at a temperature of 5°C, by the restriction 15, and reaches a state represented <BR> <BR> <BR> <BR> by a point h. The refrigerant at the point h flows as a mixture of the refrigerant liquid and the vapor at a temperature of 5°C into the evaporator 1, where the refrigerant removes heat from the process air to thus be evaporated into a saturated vapor at the state indicated by the point a on the Mollier diagram. The evaporated vapor is drawn again by the compressor 4, and thus the above cycle is repeated.

In the same manner as described above, the refrigerant flowing into the refrigerant path 42 passes through the restriction 11, the evaporating sections, the condensing sections, and the restriction 14. The refrigerant goes through states represented by points j, il, kl, i2, and k2 and reaches the a state represented by a point 1. The refrigerant flowing into the refrigerant path 44 passes through the restriction 13, the evaporating sections, the condensing sections, and the restriction 16. The refrigerant goes through states represented by points m, nl, ol, n2, and o2 and reaches a state represented by a point P-

In the heat exchanger 2, as described above, the refrigerant goes through changes of the evaporated state from the point e to the point fl or from the point gl to the point f2 in the evaporating section 51, and goes through changes of the condensed state from the point fl to the point gl or from the point f2 to the point g2 in the condensing section 52. Since the refrigerant transfers heat by way of evaporation and condensation, the rate of heat transfer is very high and the efficiency of heat exchanger is high.

In the vapor compression type heat pump HP1 including the compressor 4, the condenser 5, the restrictions 11-16, and the evaporator 1, when the heat exchanger 2 is not provided, the refrigerant at the state represented by the point d in the condenser 5 is returned to the evaporator 1 through the restrictions. Therefore, the enthalpy difference that can be used by the evaporator 1 is only 395.1-237.4 = 157.7 kJ/kg. With the heat pump HP1 according to the present embodiment which has the heat exchanger 2, however, the enthalpy difference that can be used by the evaporator 1 is 395.1-209.5 = 185.6 kJ/kg. Thus, the amount of refrigerant that is circulated to the compressor under the same cooling load and the required power can be reduced by 15 % (= 1- 157.7/185.6). Consequently, the heat pump HP1 according to the present embodiment can perform the same operation as with a well-known subcooled cycle.

FIG. 6 is a psychrometric chart showing an air- conditioning cycle in the dehumidifying air-conditioning apparatus shown in FIG. 2. In FIG. 6, the alphabetical letters K, X, L, M correspond to the encircled letters in FIG. 2.

In FIG. 6, the process air (in a state K) from the air-conditioned space 100 flows through the path 30 into the first heat exchanging portion 21 in the heat exchanger 2,

where the process air is cooled to a certain extent by the refrigerant that is evaporated in the evaporating section 51. This process can be referred to as precooling because the process air is preliminarily cooled before being cooled to a temperature lower than its dew point by the evaporator 1. While the process air is being precooled in the evaporating section 51, a certain amount of moisture is removed from the air to lower the absolute humidity of the air, and then air reaches a point X on the saturation curve.

Alternatively, the air may be precooled to an intermediate point between the point K and the point X. Further, the air may be precooled to a point that is shifted beyond the point X slightly toward a lower humidity on the saturation curve.

The process air precooled by the first heat exchanging portion 21 is introduced through the path 31 into the evaporator 1, where the air is cooled to a temperature lower than its dew point by the refrigerant which has been depressurized by the restrictions 14-16 and is evaporated at a low temperature. Moisture is removed from the air to lower the absolute humidity and the dry bulb temperature of the air, and the air reaches a point L. Although the thick line representing a change from the point X to the point L is plotted as being remote from the saturation curve for illustrative purpose in FIG. 6, it is actually aligned with the saturation curve.

The process air in the state represented by the point L flows through the path 32 into the second heat exchanging portion 22 in the heat exchanger 2, where the process air is heated, with the constant absolute humidity, by the refrigerant condensed in the condensing section 52, and reaches a point M. The process air in the point M has a sufficiently lower absolute humidity than the process air in the point K, a dry bulb temperature which is not excessively

lower than the process air in the point K, and a suitable relative humidity. The process air in the point M is then drawn by the air blower 3 and returned to the air-conditioned space 100 through the path 34.

In the air cycle on the psychrometric chart shown in FIG. 6, the amount of heat which has precooled the process air in the first heat exchanging portion 21, i. e., the amount AH of heat which has reheated the process air in the second heat exchanging portion 22, represents the amount of heat recovered, and the amount of heat which has cooled the process air in the evaporator 1 is represented by AQ. The cooling effect for cooling the air-conditioned space 100 is represented by Ai.

As described above, in the heat exchanger 2, the process air is precooled by evaporation of the refrigerant in the evaporating section 51, and the process air is reheated by condensation of the refrigerant in the condensing section 52. The refrigerant evaporated in the evaporating section 51 is condensed in the condensing section 52. The same refrigerant is thus evaporated and condensed to perform a heat exchange indirectly between the process air before being cooled in the evaporator 1 and the process air after being cooled in the evaporator 1.

In the embodiment described above, the same refrigerant is used as a heat transfer medium in the evaporator for cooling the process air to a temperature lower than its dew point, the precooler for precooling the process air, and the reheater for reheating the process air. Therefore, the refrigerant system is simplified. The refrigerant is positively circulated because the pressure difference between the evaporator and the condenser can be utilized. Since a boiling phenomenon with a phase change is applied to heat exchanges for precooling and reheating the process air, a

high heat transfer efficiency can be achieved.

In the embodiment described above, the refrigerant path is branched into the three branched refrigerant paths.

However, the present invention is not limited to three branched refrigerant paths. The refrigerant path may be branched into any number of branched refrigerant paths. FIG.

7 is a graph showing the relationship between the number of branched refrigerant paths and the temperature efficiency in a dehumidifying air-conditioning apparatus according to the present invention. It is inferred from FIG. 7 that the temperature efficiency can be improved when the number of branched refrigerant paths is larger. Thus, when a plurality of branched refrigerant paths are provided, the operative temperature of the refrigerant can gradually be changed to achieve a high efficiency of heat exchange.

A dehumidifying air-conditioning apparatus according to a second embodiment of the present invention will be described below with reference to FIGS. 8 and 9. FIG. 8 is a flow diagram of a dehumidifying air-conditioning apparatus according to the second embodiment of the present invention, and FIG. 9 is a Mollier diagram of a heat pump HP2 included inthe dehumidifying air-conditioning apparatus shown in FIG.

8. In FIGS. 8 and 9, like parts and components are denoted by the same reference numerals and characters as those of the first embodiment and will not be described below.

In the present embodiment, a refrigerant path is branched into a plurality of refrigerant paths at the downstream side of the condenser 5 to form branched refrigerant paths 142,143,144. The present embodiment differs from the first embodiment in that these branched refrigerant paths 142,143,144 extend to the interior of an evaporator 101, respectively, and joined to each other at the downstream side of the evaporator 101. Among these

branched refrigerant paths 142,143,144, the refrigerant path for the refrigerant which exchanges heat with the process air having a high temperature, i. e., the branched refrigerant path 142, has an ejector 8 provided thereon for pressurizing the refrigerant which exchanges heat with the process air having a low temperature, i. e., the refrigerant that has passed through the refrigerant path 144.

In FIG. 9, a point a represents a state of the refrigerant which has been evaporated by the evaporator 101 shown in FIG. 8, and the refrigerant is in the form of a <BR> <BR> <BR> <BR> saturatedvapor. Therefrigeranthasapressureof0. 262MPa, a temperature of 8°C, and an enthalpy of 396. 8 kJ/kg. A point b represents a state of the vapor drawn and compressed by the compressor 4, i. e., a state at the outlet port of the compressor 4. In the point b, the refrigerant has a pressure of 0.706 MPa and is in the form of a superheated vapor.

The refrigerant vapor is cooled in the condenser 5 and reaches a state represented by a point c in the Mollier diagram.

In the point c, the refrigerant is in the form of a saturated vapor and has a pressure of 0.706 MPa and a temperature of 38°C. Under this pressure, the refrigerant is cooled and condensed to reach a state represented by a point d. In the point d, the refrigerant is in the form of a saturated liquid and has the same pressure and temperature as those in the point c. The saturated liquid has an enthalpy of 237. 4 kJ/kg.

The refrigerant liquid is depressurized by the restriction 12 and reaches a state represented by a point e on the Mollier diagram. The pressure of the refrigerant liquid is an intermediate pressure between the condensing pressure in the condenser 5 and the evaporating pressure in the evaporator 101, i. e., is of an intermediate value between 0.2 62 MPa and 0.706 MPa in the present embodiment. Then, the refrigerant flows alternately through the evaporating

sections in the first heat exchanging portion 21 and the condensing sections in the second heat exchanging portion 22 and goes through states represented by points fl, gl, f2, and g2. Thereafter, the refrigerant is depressurized by the restriction 15 to a saturation pressure of 0.262 MPa at a temperature of 8°C and reaches a state represented by the point h. The refrigerant at the point h is delivered as a mixture of the refrigerant liquid and the vapor at a temperature of 8 ° C to the evaporator 101, in which the mixture removes heat from the process air and is evaporated to reach a state of the saturated vapor, which is represented by the point a in the Mollier diagram. The saturated vapor is drawn into the compressor 4 again, and the above cycle is repeated.

The refrigerant flowing into the refrigerant path 142 passes through the restriction 11, the evaporating sections, the condensing sections, and the restriction 14. The refrigerant goes through states represented by points j, il, kl, i2, and k2 and reaches the a state represented by a point 1. The refrigerant in the state represented by the point 1 flows into the evaporator 101, where the refrigerant removes heat from the process air to be evaporated and reaches a state indicated by the point q on the Mollier diagram. The refrigerant flowing into the refrigerant path 144 passes through the restriction 13, the evaporating sections, the condensing sections, and the restriction 16. The refrigerant goes through states represented by points m, nl, ol, n2, and o2 and reaches a state represented by a point P.

The refrigerant in the state represented by the point p flows into the evaporator 101, where the refrigerant removes heat from the process air to be evaporated and reaches a state indicated by the point r on the Mollier diagram. The refrigerant in the state represented by the point r is

pressurized by the ejector 8 provided on the refrigerant path 142. Specifically, in the ejector 8, the refrigerant at a low pressure in the state represented by the point r is pressurized by the refrigerant at a high pressure in the state represented by the point q. As a result, the refrigerant in the state represented by the point r and the refrigerant in the state represented by the point q reach a state of the saturated vapor, which is represented by the point a in the Mollier diagram. In this manner, with the ejector 8, since the operative temperature of the evaporator is increased to improve the theoretical cooling effect, the theoretical work of compression is reduced to achieve a high efficiency.

Further, the specific volume of the refrigerant is reduced to increase the flow rate of the refrigerant drawn by the compressor. Therefore, an amount of moisture removal is increased according to the improved cooling effect, and hence a high efficiency can be achieved.

In the vapor compression type heat pump HP2 including the compressor 4, the condenser 5, the restrictions 11-16, and the evaporator 101, when the heat exchanger 2 is not provided, the refrigerant at the state represented by the point d in the condenser 5 is returned to the evaporator 101 through the restrictions. Therefore, the enthalpy difference that can be used by the evaporator 101 is only 396.8-237.4 = 159.4 kJ/kg. With the heat pump HP2 according to the present embodiment which has the heat exchanger 2, however, the enthalpy difference that can be used by the refrigerant evaporator 101 is 396.8-209.5 = 187.3 kJ/kg.

Thus, the amount of refrigerant that is circulated to the compressor under the same cooling load and the required power can be reduced by 15 % (= 1-159. 4/187.3). Consequently, the heat pump HP2 according to the present embodiment can perform the same operation as with a well-known subcooled cycle.

While the present invention has been described in detail with reference to the preferred embodiments thereof, it would be apparent to those skilled in the art that many modifications and variations may be made therein without departing from the spirit and scope of the present invention.

For example, the number of the evaporating sections on the respective branched refrigerant paths in the first heat exchanging portion and the number of the condensing sections on the respective branched refrigerant paths in the second heat exchanging portion are not limited to the illustrated examples. With respect to the order of the refrigerant paths in the heat exchanger, the refrigerant may be introduced into the heat exchanger from the second heat exchanging portion in place of the first heat exchanging portion. In this case, the second heat exchanging portion, the first heat exchanging portion, and the second heat exchanging portion are arranged in the order named, so that the number of paths can be increased. Further, the dehumidifying air-conditioning apparatus according to the above embodiments has been described as the dehumidifying air-conditioning apparatus for air-conditioning a space. However, the dehumidifying air-conditioning apparatus according to the present invention is applicable not only to the air-conditioned space, but also to other spaces that need to be dehumidified.

Industrial Applicability The present invention is suitable for use in a heat pump with a high coefficient of performance (COP) and a dehumidifying air-conditioning apparatus which has such a heat pump and a high moisture removal per energy consumption.