Background Art FIG. 8 is a flow diagram of a conventional air- conditioning system. As shown in FIG. 8, 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. 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. 9 is a Mollier diagram in the case where HFC134a is used as the refrigerant in the conventional dehumidifying

air-conditioning apparatus. In FIG. 9, 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. 10 is a psychrometric chart showing an air- conditioning cycle in the conventional dehumidifying air- conditioning apparatus. In FIG. 10, the alphabetical letters K, L, M correspond to states in paths indicated by the encircled letters in FIG. 8. As shown in FIG. 10, 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 30% 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 object, according to an 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 first heat exchanging portion disposed in a 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 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; and a low-temperature heat source fluid path connecting the first heat exchanging portion, the evaporator, the second heat exchanging portion in the order named; wherein the refrigerant path extends alternately through the first heat exchanging portion and the second heat exchanging portion to form evaporating sections in the first heat exchanging portion and condensing sections in the second heat exchanging portion ; and the last condensing section in the second heat exchanging portion has a larger heat transfer area than the last evaporating section in the first heat exchanging portion.

According to another 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 process air; an evaporator for evaporating the refrigerant to cool the process air to a temperature lower than its dew point; a first heat exchanging portion disposed in a 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 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; and a process air path connecting the first heat exchanging portion, the evaporator, the second heat exchanging portion in the order named; wherein the refrigerant path extends alternately through the first heat exchanging portion and the second heat exchanging portion to form evaporating sections in the first heat exchanging portion and condensing sections in the second heat exchanging portion; and the last condensing section in the second heat exchanging portion has a larger heat transfer area than the last evaporating section in the first heat exchanging portion.

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 fluid 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, since the last condensing section in the second heat exchanging portion has a larger heat transfer area than

the last evaporating section in the first heat exchanging portion, the refrigerant is subcooled into a subcooled liquid in the last condensing section. Accordingly, the enthalpy difference that can be used by the evaporator can be made larger, so that the cooling effect is improved to achieve a high moisture removal (the dehumidifying performance).

Brief Description of Drawings FIG. 1 is a schematic view showing a whole arrangement of an air-conditioning system according to an embodiment of the present invention; FIG. 2 is a flow diagram of a dehumidifying air- conditioning apparatus according to an embodiment of the present invention; FIG. 3 is an enlarged view showing a refrigerant path in a heat exchanger of the dehumidifying air-conditioning apparatus shown in FIG. 2; FIG. 4A is a plan view showing a serpentine heat exchanger suitable for use in the heat exchanger of the dehumidifying air-conditioning apparatus shown in FIG. 2; FIG. 4B is a cross-sectional view taken along a line A-A of FIG. 4A; 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 flow diagram of a dehumidifying air- conditioning apparatus according to another embodiment of the present invention; FIG. 8 is a flow diagram of a conventional dehumidifying air-conditioning apparatus; FIG. 9 is a Mollier diagram of a heat pump included in

the conventional dehumidifying air-conditioning apparatus; and FIG. 10 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 an 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 an embodiment of the present invention, and FIG.

2 is a flow diagram of a dehumidifying air-conditioning apparatus according to the present embodiment. The dehumidifying air-conditioning apparatus in the present embodiment serves to cool air (process air) RA to a temperature lower than its dew point for dehumidifying the air. The dehumidifying air-conditioning apparatus includes a heat pump HP1 therein. The process air RA is lowered in humidity and supplied as the process air SA to an air-conditioned space 100 by the dehumidifying air-conditioning apparatus, for thereby maintaining a comfortable environment in the air-conditioned space 100.

As shown in FIG. 1, the dehumidifying air-conditioning apparatus mainly comprises an indoor unit 10 installed inside of the air-conditioned space 100 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 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.

As shown in FIG. 2, refrigerant paths, which are paths for circulating the refrigerant, include a path 40 connecting the evaporator 1 and the compressor 4, a path 41 connecting the compressor 4 and the condenser 5, a path 42 connecting the condenser 5 and heat exchanger 2, and a path 43 connecting the heat exchanger 2 and the evaporator 1. The refrigerant path in the heat exchanger 2 alternately extends through the first heat exchanging portion 21 and the second heat exchanging portion 22, respectively. An evaporating section 61 for evaporating the refrigerant to cool the air K 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 62 for condensing the refrigerant to heat the air L which flows through the second heat exchanging portion 22 is provided in the second heat exchanging portion 22 of the heat exchanger 2. A restriction 50 is disposed on the refrigerant path 42 at the upstream side of the first heat exchanging portion 21. A restriction 51 is disposed on the refrigerant path 43 at the downstream side of the second heat exchanging portion 22. The restrictions 50,51 may comprise orifices, capillary tubes, expansion valves, or the like.

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.

FIG. 3 is an enlarged view showing the refrigerant paths in the heat exchanger 2 of the dehumidifying air-conditioning apparatus shown in FIG. 2. The refrigerant path including the evaporating section 61 and the condensing section 62 extends through 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 in the heat exchanger 2 has an evaporating section 61a, a condensing section 62a, a condensing section 62b, an evaporating section 61b, an evaporating section 61c, a condensing section 62c, a condensing section 62d, an evaporating section 61d, an evaporating section 61e, and a condensing section 62e. The refrigerant path in the heat exchanger 2 has a condensing section 62f and a condensing section 62g subsequently to the condensing section 62e in the second heat exchanging portion 22. In the present embodiment,

the last condensing sections (62e through 62g) in the second heat exchanging portion 22 have a larger heat transfer area than the last evaporating sections (61d and 61e) in the first heat exchanging portion 21.

A serpentine heat exchanger can be used as the heat exchanger 2 in the present embodiment. FIG. 4A is a plan view showing a serpentine heat exchanger suitable for use in the heat exchanger 2 of the dehumidifying air-conditioning apparatus shown in FIG. 2. FIG. 4B is a cross-sectional view taken along a line A-A of FIG. 4A. As shown in FIGS. 4A and 4B, the first heat exchanging portion 21 through which the air K flows into the evaporator 1 and the second heat exchanging portion 22 through which the air L flows out of the evaporator 1 form respective separate spaces, each in the form of a rectangular parallelepiped. The first heat exchanging portion 21 and the second heat exchanging portion 22 have a plurality of substantially parallel heat exchange tubes 70 as refrigerant passages in planes which lie perpendicularly to the flow of the process air. While the refrigerant path in the heat exchanging portions is shown in a simplified form for illustrative purpose in FIGS. 2 and 3, the refrigerant path typically have more rows of the refrigerant passages in the heat exchanging portions with the heat exchange tubes 70.

As shown in FIG. 4B, the heat exchange tube 70 comprises a flat tube having a plurality of passages therein. Such a heat exchange tube is formed by extrusion of aluminum. A plurality of fins made of aluminum are provided between the rows of the heat exchange tube 70. As shown in FIG. 4A, the heat exchange tubes 70 form 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 depressurized by the restriction 50 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 in the first heat exchanging portion 21, 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.

The refrigerant path in the heat exchanger 2 is constructed as a continuous tube. Therefore, 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 and the condensing section 62d to heat (reheat) the process air. In this manner, the refrigerant flows through the refrigerant path in the heat exchanger while changing in phase between the vapor phase and the liquid phase. Thus, heat is exchanged indirectly between the process air before being cooled by the evaporator 1 and the process air which has been cooled by the evaporator 1 and lowed in absolute humidity.

As described above, the last condensing sections (62e through 62g) in the second heat exchanging portion 22 have a larger heat transfer area than the last evaporating sections (61d and 61e) in the first heat exchanging portion 21.

Therefore, the refrigerant is subcooled into a subcooled liquid in the last condensing sections 62e through 62g. The refrigerant liquid subcooled in these condensing sections is depressurized and expanded by the restriction 51 provided at

the downstream side of the second heat exchanging portion 22, for thereby lowering its pressure. Then, the refrigerant enters the evaporator 1 and is evaporated therein. The refrigerant cools the process air flowing through the first heat exchanging portion 21, 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 present embodiment 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 saturated vapor.

The refrigerant has a pressure of 0.350 MPa, a temperature of 5°C, and an enthalpy of 401.5 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 0.963 MPa and is in the form of a superheated vapor.

The refrigerant vapor at the point b is cooled in the

condenser 2 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 0.963 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 253.4 kJ/kg.

The refrigerant liquid is depressurized by the restriction 50 and flows into the evaporating section 61a 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 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 0.963 MPa and 0.350 MPa in the present embodiment.

In the evaporating section 61a, 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 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 62a, 62b. In the condensing sections 62a, 62b, 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 at the point gl is evaporated in the

evaporating sections 61b, 61c and then condensed in the condensing sections 62c, 62d. The evaporation in the evaporating sections 61b, 61c and the condensation in the condensing sections 62c, 62d are not shown for illustrative purpose in the Mollier diagram shown in FIG. 5. Thereafter, the refrigerant is evaporated in the evaporating sections 61d, 61e and reaches a state represented by a point f2.

The refrigerant at the poring f2 flows into the condensing section 61e and condensed therein while increasing the liquid phase thereof. Then, the refrigerant is condensed and subcooled in the condensing sections 62f, 62g and reaches a state represented by a point g2. In the point g2, the refrigerant is in the form of a subcooled liquid, and has a temperature of 14°C and an enthalpy of 219.1 kJ/kg.

The refrigerant liquid at the point g2 is depressurized to 0.350 MPa, which is a saturated pressure at a temperature of 5°C, by the restriction 51, and reaches a state represented 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 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 61, 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 62. 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 50,51, 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 401.5-253.4 = 148.1 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 401.5-219.1 = 182.4 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 19 % (= 1-148.1/182.4). Consequently, the heat pump HP1 according to the present embodiment can perform the same operation as with a well-known subcooled cycle.

As described above, the last condensing sections (62e through 62g) in the second heat exchanging portion 22 have a larger heat transfer area than the last evaporating sections (61d and 61e) in the first heat exchanging portion 21.

Therefore, the refrigerant is subcooled into a subcooled liquid in the last condensing sections 62e through 62g. Accordingly, the enthalpy difference that can be used by the evaporator 1 can be made larger, so that the cooling effect is improved to achieve a high moisture removal (the dehumidifying performance).

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 states in the paths indicated by 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 61. 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 61, 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 restriction 51 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 62, 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 61, and the process air is reheated by condensation of the refrigerant in the condensing section 62.

The refrigerant evaporated in the evaporating section 61 is condensed in the condensing section 62. The same refrigerant is thus evaporated and condensed to perform 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 present embodiment, the condenser heats the outside air OA as cooling air. However, the air that has been heated by the second heat exchanging portion may further be heated (reheated) by the condenser. FIG. 7 shows an example in which the air that has been heated by the second heat exchanging portion 22 is heated (reheated) by the condenser 5 and supplied to the air-conditioned space 100 in the dehumidifying air-conditioning apparatus of the above embodiment.

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 refrigerant path in the first heat exchanging portion and the number of the condensing sections on the refrigerant path in the second heat exchanging portion are not limited to the illustrated examples.

Further, the dehumidifying air-conditioning apparatus according to the above embodiment 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.