JP2000249366A - Heat pump, dehumidifier and dehumidifying method employing it - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a heat pump, a dehumidifier/air conditioner and a dehumidifying method having high COP. SOLUTION: The heat pump comprises a refrigerant compressor 260, a first heat exchanger 220 for robbing heat from a compressed refrigerant through a high temperature fluid B and condensing the refrigerant under a first pressure, a first throttle 360 for reducing the pressure of condensed refrigerant to a second pressure, a second heat exchanger 300 for evaporating the refrigerant subjected to pressure reduction through the first throttle 360 with heat from a first fluid A under the second pressure and condensing a second fluid C by robbing heat therefrom, a second throttle 270 for reducing the pressure of condensed refrigerant to a third pressure, a third heat exchanger 210 for evaporating the refrigerant subjected to pressure reduction through the second throttle 270 by imparting heat from the low temperature fluid A under the third pressure, and a gas/liquid separator 350 for separating the refrigerant subjected to pressure reduction to the second pressure to refrigerant gas and refrigerant gas between the first throttle 360 and the second heat exchanger 300. The refrigerant compressor 260 has an intermediate port 260a for sucking gas at an intermediate pressure between first and third pressures and the refrigerant gas separated by the gas/liquid separator 350 is sucked through the intermediate port 260a.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a heat pump, a dehumidifying apparatus and a dehumidifying method using the same, and more particularly to a compression heat pump having an economizer using a heat exchanger, and a dehumidifying apparatus and a dehumidifying method using such a compressed heat pump. It is.

[0002]

2. Description of the Related Art As shown in FIG. 10, there has been a so-called desiccant air conditioning system using a heat pump as a heat source. In the air conditioning system of FIG. 10, as a heat pump,
A compression heat pump HP using the compressor 260 is used. This air conditioning system has a desiccant rotor 103
A path of the processing air A for adsorbing the water by the heat source, and a path of the regeneration air B for desorbing and regenerating the moisture in the desiccant after passing through the desiccant rotor 103 after being heated by the heating source and adsorbing the moisture. An air conditioner having a sensible heat exchanger 104 between the treated air to which moisture has been adsorbed and the regenerated air before regenerating the desiccant (desiccant) of the desiccant rotor 103 and before being heated by the heating source; A heat pump HP, and using the high heat source of the compression heat pump HP as a heating source to supply the regenerated air of the air conditioner to a heater 22.
At 0, the desiccant is regenerated, and the cooler 210 cools the processing air of the air conditioner using the low heat source of the compression heat pump as a cooling heat source.

In this air conditioning system, since the compression heat pump HP is configured to simultaneously cool the processing air of the desiccant air conditioner and heat the regeneration air, the compression heat pump HP is driven by driving heat externally applied to the compression heat pump HP. Generates a cooling effect of the processing air, and furthermore, the desiccant can be regenerated with the heat obtained by summing the heat pumped from the processing air by the heat pump action and the driving heat of the compression heat pump HP. As a result, a high energy saving effect can be obtained. Further, a heat exchanger 121 between the regenerated air between the sensible heat exchanger 104 and the heater 220 and the regenerated air exiting the desiccant rotor 103 is provided to further enhance the energy saving effect.

The operation of the desiccant air conditioner shown in FIG. 10 will now be described with reference to a psychrometric chart shown in FIG.
In FIG. 11, the state of air is indicated by alphabets K to P and Q to X. This symbol corresponds to the alphabet circled in the flowchart of FIG.

[0005] In FIG. 11, the processing air (state K) from the air-conditioned space 101 is desiccant adsorbed by the desiccant rotor 103 to lower the absolute humidity, while the desiccant heat of adsorption raises the dry bulb temperature to the state L.
Then, in the sensible heat exchanger 104, the air is cooled in the state M while keeping the absolute humidity constant, and enters the cooler 210. Here, the air is further cooled at a constant absolute humidity to become air in the state N, and humidified by the humidifier 105 to lower the dry-bulb temperature to become air in the state P, and is returned to the air-conditioned space 101. On the other hand, the outside air in the state Q is sent to the sensible heat exchanger 104, where it cools the processing air to be heated to the state R, enters the heat exchanger 121, and is further heated to the state S. Then, the desiccant is heated by the heater 220 to reach the state T, and the desiccant is regenerated by the desiccant rotor 103. The desiccant itself has a high absolute humidity, the dry bulb temperature decreases, and the state becomes U, and the regenerated air is heated by the heat exchanger 121. As a result, the temperature of the dry bulb itself is lowered, and the air becomes the state V and is exhausted EX.

Further, the operation of the compression heat pump HP shown in FIG. 10 will be described with reference to the Mollier diagram of FIG. FIG. 12 is a Mollier diagram of the refrigerant HFC134a. Point a indicates the state of the refrigerant evaporated in the cooler 210, and is in the state of a saturated gas. The pressure is 4.2kg / cm
^{2. The} temperature is 10 ° C, the enthalpy is 148.83 kcal
/ Kg. The state where the gas is sucked and compressed by the compressor 260 and the state at the discharge port of the compressor 260 are indicated by a point b. In this state, the pressure is 19.3 kg / cm ² , the temperature is 78 ° C., and the state is a superheated gas. This refrigerant gas is cooled in the heater (condenser as viewed from the refrigerant side) 220 and reaches a point c on the Mollier diagram. This point is a state of a saturated gas, the pressure is 19.3 kg / cm ² , and the temperature is 65
° C. Further cooling and condensing under this pressure, the point d
To reach. This point is a state of the saturated liquid, and the pressure and the temperature are the same as those of the point c, the pressure is 19.3 kg / cm ² , and the temperature is 65
° C, and the enthalpy is 122.97 kcal / kg. This refrigerant liquid is decompressed by the expansion valve 270 and has a temperature of 1
The pressure is reduced to 4.2 kg / cm ^2, which is a saturation pressure of 0 ° C., and reaches a cooler (evaporator as viewed from the refrigerant) 210 as a mixture of a refrigerant liquid and a gas at 10 ° C. where heat is taken from the processing air. , Evaporates to become a saturated gas in the state of the point a on the Mollier diagram, is sucked into the compressor 260 again, and the above cycle is repeated.

[0007]

According to the conventional heat pump described above, the enthalpy difference indicating the cooling effect of the refrigerant per unit weight is determined by the enthalpy of the saturated gas line at the evaporation pressure of the refrigerant and the enthalpy of the saturated gas at the condensation pressure. The difference from the enthalpy of the line, that is, 148.83-12 in the example of FIG.
2.97 = 25.86 kcal / kg, which is not necessarily large, so that the coefficient of performance C of the compression heat pump HP
The OP had to be low. Also, desiccant air conditioners (dehumidifiers) using such heat pumps have had to reduce COP.

Accordingly, an object of the present invention is to provide a heat pump having a high COP, a dehumidifier using the same, and a dehumidifying method.

[0009]

In order to achieve the above object, a heat pump according to the first aspect of the present invention comprises a refrigerant compressor 26 for compressing a refrigerant as shown in FIG.
0; a first heat exchanger 220 for removing heat from the refrigerant compressed by the refrigerant compressor 260 by the high-temperature fluid B and condensing the refrigerant under a first pressure; and a first heat exchanger 220 for condensing the refrigerant. A first throttle 360 for reducing the pressure of the cooled refrigerant to a second pressure;
Under the second pressure, the refrigerant decompressed by the first throttle 360 is evaporated by the heat from the first fluid A, and after the evaporation, the refrigerant is deprived of heat by the second fluid C to remove the refrigerant. A second heat exchanger 300 for condensing the refrigerant; a second restrictor 270 for condensing the refrigerant in the second heat exchanger 300 and then depressurizing the refrigerant to a third pressure; and a cryogenic fluid under the third pressure. A third heat exchanger 210 configured to apply heat from A to evaporate the refrigerant depressurized by the second throttle 270;
A gas-liquid separator 350 for separating the refrigerant decompressed to the second pressure into a refrigerant liquid and a refrigerant gas between the first throttle 360 and the second heat exchanger 300; Machine 2
60 has an intermediate port 260a for suctioning gas at a pressure intermediate between the first pressure and the third pressure, and supplies the refrigerant gas separated by the gas-liquid separator 350 to the intermediate port 260a.
It is characterized by being constituted so that it may be inhaled. Typically, the refrigerant liquid separated by the gas-liquid separator 350 is configured to flow to the second heat exchanger 300.

[0010] With this configuration, the second heat exchanger is provided, and the second heat exchanger cools the first fluid, for example, the processing air by evaporating the refrigerant, and converts the evaporated refrigerant into the second fluid, for example. Since it is configured to cool and condense by the outside air, it is possible to use evaporation heat transfer and condensation heat transfer having a high heat transfer coefficient, so that heat transfer between the first fluid and the second fluid with a high heat transfer coefficient is achieved. Can be achieved. Further, since the heat transfer between the first fluid and the second fluid is performed via the refrigerant, the components of the heat pump can be easily arranged. In addition, since it has a gas-liquid separator,
Here, the separated refrigerant gas can be led to the intermediate port of the compressor. Therefore, the refrigerant cycle can be configured as a so-called economizer cycle.

If the low-temperature fluid is a first fluid, for example, processing air, the processing air can be cooled by the second heat exchanger and then further cooled by the third heat exchanger.

Further, the refrigerant liquid separated by the gas-liquid separator 350 is converted into a processing air cooler 300 as a second heat exchanger.
In particular, when configured to flow through the evaporating section 251 as the first fluid flow path, the refrigerant that flows into the processing air cooler 300 to cool, for example, the processing air by evaporation can be mainly composed of the liquid phase. Also, for example, the gas phase of the refrigerant condensed by the outside air can be made uniform.

Further, the third restriction 26 is provided between the gas-liquid separator 350 and the intermediate port 260a.
2 may be provided. At this time, the differential pressure between the second pressure and the suction pressure of the intermediate port can be adjusted by the third throttle.

Further, according to claim 3, for example, FIG.
As shown in the above, in the above heat pump, the second heat exchanger 300 is provided in the first section 3 in which the first fluid A flows.
10; a second section 320 through which the second fluid C flows;
A first fluid flow path 251 through which the refrigerant that exchanges heat with the first fluid A passes through the first compartment 310; and a heat exchange with the second fluid C that passes through the second compartment 320 A second fluid flow path 252 through which a refrigerant flows; the refrigerant flows from the first fluid flow path 251 to the second fluid flow path 252, and flows along the flow path of the first fluid flow path 251. On the hot side, the refrigerant evaporates under the second pressure and the second fluid flow path 25
In the second flow path side heat transfer surface, the refrigerant may be condensed substantially under the second pressure.

With this configuration, the refrigerant flows from the first fluid passage 251 to the second fluid passage 252,
The heat transfer coefficient is high because it flows in one direction as a whole.

In order to achieve the above object, a dehumidifying device according to a fourth aspect of the present invention provides a heat pump HP according to any one of the first to third aspects, as shown in FIG.
1; a moisture adsorbing device 103 that is provided upstream of the flow of the first fluid A with respect to the second heat exchanger 300 and has a desiccant that adsorbs moisture in the first fluid A.

According to this structure, a water adsorbing device which is provided on the upstream side of the flow of the first fluid with respect to the second heat exchanger and has a desiccant for adsorbing the water in the first fluid A is provided. First, water content is adsorbed by desiccant
Can be cooled by the second heat exchanger.

Further, the moisture adsorption device 103 may be provided on the downstream side of the high-temperature fluid with respect to the first heat exchanger 220 so as to desorb moisture contained in the desiccant by the high-temperature fluid. Good. At this time, the desiccant can be regenerated by the high-temperature fluid heated by the first heat exchanger.

In order to achieve the above object, a method for dehumidifying a fluid according to the present invention according to claim 5 includes, for example, a first step of condensing a refrigerant at a first pressure by a heat exchanger 220 of FIG. 1; 360, a second step in which the refrigerant condensed at the first pressure is reduced to a second pressure to evaporate a part of the refrigerant; a part of the refrigerant gas evaporated in the second step and a remaining refrigerant liquid And e.g., in a first fluid flow path 251 (FIG. 2), evaporating the separated refrigerant liquid substantially under the second pressure. A fourth step of cooling the first fluid with the evaporating refrigerant; and causing the refrigerant gas evaporated in the fourth step to pass through the second fluid flow path 252 (FIG. 2) substantially under the second pressure. A fifth step of condensing the refrigerant in the fifth step; For example, a sixth step of reducing the pressure of the refrigerant in the sixth step by the throttle 270 to a third pressure; and a seventh step of evaporating the refrigerant reduced in the sixth step by, for example, the heat exchanger 210 to cool the first fluid A. An eighth step of increasing the pressure of the refrigerant gas evaporated in the seventh step to the first pressure by, for example, a compressor 260; and increasing the pressure of the refrigerant gas separated in the third step in the eighth step. A ninth step of merging with the refrigerant gas, for example, at the intermediate port 260a of the compressor during the pressurization; a tenth step of adsorbing moisture in the first fluid with, for example, a desiccant of the moisture adsorbing device 103; An eleventh step of heating the high-temperature fluid B with the refrigerant condensed in the first step, and desorbing water from the desiccant adsorbed in the tenth step with the heated high-temperature fluid to regenerate the desiccant. . Typically, the first fluid is process air of a dehumidifying air conditioner, and the high temperature fluid is regeneration air.

With this configuration, a third step as a gas-liquid separation step is provided, and the separated refrigerant liquid is evaporated to form the first liquid.
And a fifth step of condensing the refrigerant gas evaporated in the fourth step, so that evaporation heat transfer and condensation heat transfer having a high heat transfer coefficient can be utilized. In addition, since a ninth step of combining the refrigerant gas separated in the gas-liquid separation step with the refrigerant gas pressurized in the eighth step while the pressure is rising is provided, a so-called economizer cycle can be realized.

[0021]

Embodiments of the present invention will be described below with reference to the drawings. In the drawings, the same or corresponding members are denoted by the same or similar reference numerals, and duplicate description is omitted.

Referring to FIG. 1, a compression heat pump HP1 and a heat pump HP according to a first embodiment of the present invention.
A configuration of a dehumidifying air-conditioning apparatus (air-conditioning system) according to a second embodiment of the present invention including the above-described configuration will be described. This air conditioning system lowers the humidity of processing air by a desiccant (drying agent), and supplies air-conditioned space 101 to which processing air is supplied.
Is to maintain a comfortable environment.

In the figure, a blower 102 for circulating the processing air along the path of the processing air A from the air-conditioned space 101,
Desiccant rotor 103 as a moisture adsorber filled with desiccant, processing air cooler 300 as a second heat exchanger of the present invention, refrigerant evaporator as a third heat exchanger (cooler as viewed from processing air) 210 and in this order, and is configured to return to the air-conditioned space 101.

A heat exchanger 121 for exchanging heat between the regeneration air before entering the desiccant rotor 103 and the subsequent regeneration air along a path from the outdoor OA to the regeneration air B, a refrigerant as a first heat exchanger Condenser (heater as seen from regenerated air)
220, a desiccant rotor 103, a blower 140 for circulating regeneration air, and a heat exchanger 121 are arranged in this order, and are configured to exhaust EX to the outside.

A processing air cooler 300 and a blower 160 for circulating the cooling fluid are arranged in this order along the path from the outdoor OA to the outside air C as the cooling fluid, and the exhaust air is exhausted to the outside. It is configured.

A compressor 260, a refrigerant condenser 220, a throttle 360, a gas-liquid separator 350, and a processing air for compressing the refrigerant evaporated and gasified by the refrigerant evaporator along the path of the refrigerant from the refrigerant evaporator 210. The cooler 300 and the throttle 270 are arranged in this order, and return to the refrigerant evaporator 210 again.

Here, the compressor 260 is a refrigerant evaporator 210
And an intermediate port 260a for giving an intermediate pressure between the evaporation pressure in the refrigerant condenser 220 and the condensation pressure in the refrigerant condenser 220. The gas-liquid separator 350 and the intermediate port 260a are connected by a refrigerant gas pipe 340. A throttle 262 is inserted and arranged in the refrigerant gas pipe 340 between the gas-liquid separator 350 and the intermediate port 260a. Thus, the heat pump HP1 is configured.

The desiccant rotor 103 is formed as a thick disk-shaped rotor that rotates around a rotation axis AX as shown in FIG. 13, and the rotor is filled with a desiccant with a gap through which gas can pass. ing.
For example, a large number of tubular drying elements 103a are bundled so that the central axis is parallel to the rotation axis AX. The rotor 103 is configured to rotate in one direction around the rotation axis AX. Further, the flow path of the processing air A and the flow path of the regeneration air B include a plane including the rotation axis AX, and are separated by a partition plate whose end is provided close to the surface in the thickness direction of the desiccant rotor 103. The processing air A and the regeneration air B flow in and out of the axis of rotation AX. As described above, each of the drying elements 103a is arranged so as to alternately contact the processing air A and the regeneration air B as the rotor 103 rotates.

In FIG. 13, the desiccant rotor 103
Is partially broken away. Although the figure shows a gap between the outer periphery of the desiccant rotor 103 and a part of the drying element 103a, the drying elements 103a are actually bundled and tightly packed in the entire disk.

In such an apparatus, typically, the processing air A (indicated by a white arrow in the figure) and the regeneration air B (indicated by a black solid arrow in the figure) are each circular in parallel with the rotation axis AX. Is configured to flow in a substantially half area of the desiccant rotor 103 in the counterflow type.

The desiccant may be granulated and filled in the tubular drying element 103a, the tubular drying element 103a itself may be formed of a desiccant, or the desiccant may be applied to the drying element 103a. Alternatively, the drying element 103a may be made of a porous material, and the material may contain desiccant. The drying element 103a may be formed in a cylindrical shape with a circular cross section as shown in the figure, or formed in a hexagonal cylindrical shape,
They may be bundled to form a honeycomb shape as a whole. In any case, the air is configured to flow in the thickness direction of the disk-shaped rotor 103.

Since a large amount of regeneration air must be passed through the heat exchanger 121, for example, as shown in FIG. 14, a cross-flow type in which low-temperature regeneration air B1 and high-temperature regeneration air B2 flow orthogonally. And a rotary heat exchanger filled with a heat storage material having a large heat capacity is used in place of the drying element in a structure similar to that of the heat exchanger of FIG. In this case, the processing air A in FIG.
1 corresponds to the regeneration air B and the high-temperature regeneration air B2.

Next, referring to FIG. 2, heat pump HP1
An example of the configuration of the second heat exchanger suitable for use in the second embodiment will be described. In the figure, a heat exchanger 300 includes a first section 310 through which processing air A as a first fluid flows, and a second section 320 through which outside air C, which is a cooling fluid as a second fluid, flows.
They are provided adjacent to each other with one partition 301 interposed therebetween.

A plurality of heat exchange tubes as a fluid flow path through which the refrigerant 250 flows are provided substantially horizontally through the first section 310, the second section 320, and the partition wall 301. In the heat exchange tube, a portion penetrating the first section has an evaporating section 251 as a first fluid flow path.
(A plurality of evaporating sections 251A, 251B, 251
C. Hereinafter, when it is not necessary to discuss a plurality of evaporation sections individually, the part passing through the second section is simply a condensing section 252 (a plurality of condensing sections are referred to as 252A). , 25
2B and 252C. Hereinafter, when it is not necessary to discuss a plurality of condensation sections individually, it is simply referred to as 252).

In the form of the heat exchanger shown in FIG. 2, the evaporating section 251A and the condensing section 252A are formed as an integral flow path by one tube. The same applies to the evaporating sections 251B, C and the condensing sections 252B, C. Therefore, in combination with the fact that the first section 310 and the second section 320 are provided adjacent to each other with one partition wall 301 interposed therebetween, the heat exchanger 300 can be formed to be small and compact as a whole. .

In the form of the heat exchanger of FIG. 2, the evaporating sections are arranged in the order of 251A, 251B, 251C from the top in the figure, and the condensing sections are also 252A, 252A,
52B and 252C.

On the other hand, the processing air A is configured to enter the first section in FIG. The outside air C, which is a cooling fluid, is configured to enter the second section in the drawing through the duct 171 from below and flow out from above.

Further, in the second section, the upper part thereof, above the heat exchange tube constituting the condensing section 252,
A watering pipe 325 is arranged. Watering pipe 325
Are provided with nozzles 327 at appropriate intervals to condense the water flowing in the watering pipe 325 to the condensing section 2.
It is configured to be sprayed to the heat exchange tube constituting 52 or to the outside air C.

A vaporizing humidifier 165 is provided at the entrance of the outside air C which is a cooling fluid as a second fluid in the second section. The vaporizing humidifier 165 is made of a material that has a hygroscopic property and a gas permeability, such as a ceramic paper and a nonwoven fabric.

Here, the evaporation pressure in the evaporation section 251 and, consequently, the condensation pressure in the condensation section 252,
That is, the second pressure is determined by the temperature of the processing air A and the temperature of the outside air C which is a cooling fluid. Heat exchanger 3 shown in FIG.
No. 00 utilizes the evaporation heat transfer and the condensation heat transfer, so that the heat transfer coefficient is very excellent and the heat exchange efficiency is very high.
Further, since the refrigerant flows from the evaporating section 251 to the condensing section 252, that is, is forced to flow in substantially one direction, the heat exchange efficiency between the processing air and the outside air as the cooling fluid is high.

Here, the heat exchange efficiency φ is defined as TP1, the inlet temperature of the heat exchanger of the fluid on the high temperature side, TP2, the inlet temperature of the heat exchanger of the fluid on the low temperature side is TC1, and the outlet temperature is TC2. When attention is paid to the cooling of the fluid on the high temperature side, that is, when the purpose of the heat exchange is cooling, φ = (TP1-
TP2) / (TP1-TC1), when attention is paid to heating of a low-temperature fluid, that is, when the purpose of heat exchange is heating, φ =
It is defined as (TC2-TC1) / (TP1-TC1).

Evaporation section 251, condensation section 2
It is preferable to form a high-performance heat transfer surface by forming a spiral groove such as a linear groove on the inner surface of the barrel of the rifle gun on the inner surface of the heat exchange tube constituting 52. The refrigerant liquid flowing inside usually flows so as to wet the inner surface,
When the spiral groove is formed, the heat transfer coefficient is increased because the boundary layer of the flow is disturbed.

Although the processing air flows through the first section, it is preferable that the fin mounted on the outside of the heat exchange tube is processed into a louver shape to disturb the flow of the fluid. Preferably, the fins are similarly configured to disrupt the flow of fluid when the outside air flows into the second compartment but does not spray water. However, when water is sprayed, it is preferable to further provide a corrosion-resistant coating as flat plate fins. This is to prevent corrosive substances possibly mixed in the water from condensing and condensing by evaporation to corrode the fins or tubes. The fins are preferably made of aluminum, copper, or an alloy thereof.

The operation of the dehumidifying air-conditioning apparatus according to the second embodiment will be described with reference to FIG. In FIG. 3, an alphabet symbol D,
E, K to N, and Q to X indicate the state of air in each part. This symbol corresponds to the circled alphabet in the flow diagram of FIG.

First, the flow of the processing air A will be described. In FIG. 3, the processing air (state K) from the air-conditioned space 101 is:
The air is sucked by the blower 102 through the processing air path 107 and is sent to the desiccant rotor 103 through the processing air path 108. Here, the drying element 103a
Water is adsorbed by the desiccant in (FIG. 13) to lower the absolute humidity, and the dry bulb temperature is raised by the heat of adsorption of the desiccant to reach the state L. This air is treated air path 1
09 through the first section 310 of the process air cooler 300
Where the absolute humidity is constant and evaporation section 2
Cooled by the refrigerant evaporating in 51 (FIG. 2), it becomes air in the state M, and enters the cooler 210 through the path 110.
Here, the air is further cooled at a constant absolute humidity to be in the state N. This air is dried and cooled, and is treated as a treatment air SA having an appropriate humidity and an appropriate temperature as duct 111.
And is returned to the air-conditioned space 101.

Next, the flow of the regeneration air B will be described. In FIG. 3, the regeneration air (state Q) from the outdoor OA is sucked through the regeneration air path 124 and sent to the heat exchanger 121. Here, heat exchange is performed with the high-temperature regenerated air to be exhausted (air in state U described later) to raise the dry-bulb temperature to become air in state R. This air is sent to a refrigerant condenser (heater as viewed from the regeneration air) 220 through a path 126, where it is heated to increase the dry-bulb temperature and become air in state T. This air is sent through a path 127 to the desiccant rotor 103 where the drying element 10
The desiccant in 3a (FIG. 13) deprives the desiccant of water, that is, desorbs and regenerates the water, thereby increasing the absolute humidity and lowering the dry-bulb temperature by the heat of desiccant water desorption to reach the state U. This air is sucked through a passage 128 into a blower 140 for circulating the regeneration air, sent through a passage 129 to the heat exchanger 121, and as described above, before being sent to the desiccant rotor 103 (state). The air itself exchanges heat with the air (Q air) to lower the temperature to become air in the state V, and is exhausted EX through the passage 130.

Next, the flow of the outside air C which is a cooling fluid as the second fluid will be described. Outside air C (state Q) is outdoor OA
Through the path 171 to the second section 320 of the process air cooler 300. Here, first, the humidifier 165
Absorbs water, changes the enthalpy, raises the absolute humidity, and lowers the dry-bulb temperature to form air in state D.
State D is almost on the saturation line of the wet vapor diagram. This air cools the refrigerant in the condensing section 252 while absorbing the water sprayed and supplied by the watering pipe 325 in the second compartment 320. This air rises in absolute humidity and dry-bulb temperature substantially along the saturation line, becomes air in state E, and is exhausted EX through the passage 172 by the blower 160 provided in the middle of the passage 172.

Referring now further to FIG. 3, humidifier 16
5. The operation of the watering pipe 325 will be described. In the air conditioner as described above, as can be seen from the cycle on the air side shown in the psychrometric chart of FIG. 3, the amount of heat added to the regenerated air for the desiccant regeneration of the device is pumped from the processing air by ΔH. Assuming that the heat amount is Δq and the driving energy of the compressor is Δh, ΔH = Δq + Δh. The cooling effect ΔQ obtained as a result of the regeneration based on the heat quantity ΔH increases as the temperature of the outside air (state Q) to be heat-exchanged with the treated air (state L) after the adsorption of moisture is lower. That is, it becomes larger as ΔQ−Δq becomes larger in the figure. Therefore, spraying water on the outside air as the cooling fluid is useful for enhancing the cooling effect. In FIG. 3, points indicated as a state M ′ located between the states L and M and a state N ′ located between the states M and N are points where the evaporating humidifier 165 and the sprinkling pipe 325 are illustrated. 7 conceptually shows the positions of state M and state N when not used.

Returning now to FIG. 1, the heat pump HP1
Will be described. In the figure, the refrigerant gas compressed by the refrigerant compressor 260 passes through the refrigerant gas pipe 201 connected to the discharge port of the compressor 260 and is supplied to the regeneration air heater 2.
20. The temperature of the refrigerant gas compressed by the compressor 260 is increased by the heat of compression, and the heat heats the regeneration air (described later). The refrigerant gas itself is deprived of heat and condenses.

The refrigerant outlet of the heater 220 is connected to the heat exchanger 30
The refrigerant passage 202 is connected to the inlets of the evaporating sections 251A, 251A, B, and C through a refrigerant path 202. A throttle 360 such as an expansion valve is provided in the middle of the refrigerant path 202.
A gas-liquid separator 350 is provided between 60 and the evaporation sections 251A, 251B, 251C.

The liquid refrigerant discharged from the heater (cooler or condenser as viewed from the refrigerant side) 220 at a first pressure is reduced to a second pressure by an expansion valve 360 as a first throttle, and expands. Some liquid refrigerant evaporates (flashes). The refrigerant in which the liquid and the gas are mixed is supplied to the gas-liquid separator 350.
The refrigerant liquid is separated into a refrigerant liquid and a refrigerant gas, and the refrigerant liquid reaches the evaporating sections 251A, B, and C, and the refrigerant evaporates in the tubes of the evaporating section to cool the processing air flowing through the first section.

Evaporation section 251 and condensation section 2
52 is a series of tubes, ie, configured as an integral flow path, so that the evaporated refrigerant gas (and the non-evaporated refrigerant liquid) flows into the condensing section 252 and passes through the second section. The heat is deprived by the flowing outside air and the sprayed water and condensed.

The outlet side of the condensing section 252 is connected to a header that collects the condensing sections 252A, B, and C. The header is connected to an expansion valve 270 as a second restrictor by a refrigerant liquid pipe 203 and further to a refrigerant pipe 204. It is connected to a cooler 210. The refrigerant liquid condensed under the second pressure in the condensing section 252 is reduced to a third pressure by the throttle 270 and expanded to lower the temperature, enters the cooler 210 and evaporates. Alternatively, the processing air as a low-temperature fluid is cooled. As the throttles 360 and 270, for example, an orifice or a capillary tube other than the expansion valve may be used.

Cooler (evaporator as viewed from refrigerant side) 210
The refrigerant evaporated and gasified by the above is guided to the suction side of the refrigerant compressor 260, and the above cycle is repeated.

The gas-liquid separator 350 includes a container into which a mixture of gas and liquid flows, and a baffle plate 355 disposed in the container so as to face the inlet of the gas-liquid mixture. I have. The gas-liquid mixture collides with the baffle plate 355 and the liquid is separated from the gas, and the gas flows out of the gas outlet provided in the vessel alongside the gas-liquid mixture inlet, and is connected to the gas outlet. The refrigerant flows through the refrigerant pipe 340 to the intermediate port 260a of the compressor 260. An eliminator (not shown) may be provided near the refrigerant gas outlet of the gas-liquid separator 350. The refrigerant liquid flows out from a liquid outlet provided below the container of the gas-liquid separator 350 in the vertical direction. Refrigerant pipes 430A, 430B, and 430C are connected to the liquid outlets, and communicate with the evaporation sections 251A, 251B, and 251C, respectively.

The pressure of the intermediate port 260a is determined so as to be a pressure corresponding to the second pressure mainly determined by the type of the refrigerant, the temperature of the first fluid, and the temperature of the second fluid. That is, the intermediate port 260a in the compressor 260
Determine the position of If these two pressures are almost the same,
The third stop 262 shown in FIG. 1 is unnecessary.

However, in general, the temperature of the processing air as the first fluid must be changed according to the air-conditioning load of the air-conditioned space 101 and the preference of humans therein.
The temperature of the outside air as a fluid varies depending on the season and the time of day. Therefore, in order to flexibly respond to these changes, the suction pressure of the intermediate port 260a is designed to be equal to or slightly lower than the lowest pressure considered as the second pressure. The third throttle 262 is used to compensate for the pressure difference. The throttle 262 is shown as an orifice in the drawing, but may be a manual valve or an automatic control valve capable of adjusting the degree of throttle. In the case of an automatic control valve, it may be operated by receiving a signal from a controller (not shown) for adjusting the second pressure to a set value, for example.

Referring to the Mollier diagram of FIG. 4, the operation of the heat pump HP1 according to the first embodiment of the present invention in the air conditioning system of FIG. 1 will be further described. FIG. 4 shows the refrigerant HF
It is a Mollier diagram in case C134a is used. In this diagram, the horizontal axis is enthalpy and the vertical axis is pressure.

In the figure, the point a is the state of the refrigerant outlet of the refrigerant evaporator 210 of FIG. 1 and is in the state of saturated gas. The evaporation pressure as the first pressure is 4.2 kg / cm ² , and the temperature is 10
° C, enthalpy is 148.83 kcal / kg.
The state where this gas is sucked by the compressor 260 and compressed to the second pressure, that is, the state at the pressure of the intermediate port 260a is indicated by a point k. In this state, the pressure is a saturation pressure of 40 ° C. Here, the pressure loss of the flow is neglected.
FIG. 4A shows a case where the third throttle 262 is not provided or a case where the control valve as the throttle is fully opened.

The refrigerant gas in the state at the point k and the saturated gas in the state at the point h flowing into the compressor from the intermediate port 260a as will be described later are mixed, and as a gas at the state at the point m, the compressor 2
Further compression by 60 leads to point b. At point b, the refrigerant gas has a pressure of 19.3 kg / cm ² and a temperature of about 75 ° C.
And is in a state of superheated gas.

This refrigerant gas is cooled in the refrigerant condenser 220 and reaches a point c on the Mollier diagram. Point c is a saturated gas state, the pressure is 19.3 kg / cm ² , and the temperature is 65
° C. Further cooling and condensing under this pressure, the point d
To reach. This point is a state of the saturated liquid, and the pressure and the temperature are the same as those of the point c, the pressure is 19.3 kg / cm ² , and the temperature is 65
° C, and the enthalpy is 122.97 kcal / kg.

The refrigerant liquid in the state at the point d is decompressed by the throttle 360 and flows into the gas-liquid separator 350 to reach the point s. The refrigerant gas in the state at point s is the second pressure of the present invention, the gas in the state of intersection h with the saturated gas line on the isobar at the saturation pressure corresponding to 40 ° C. and the intersection with the saturated liquid line. The liquid in the state of e is separated. The refrigerant gas in the state at the point h flows into the intermediate port 260a of the compressor 260 via the pipe 340. As described above, the refrigerant gas is mixed with the refrigerant gas in the state at the point k, and becomes a gas in the state at the point m.
The compression stroke after 0a is received.

At the point s, the ratio between the refrigerant liquid and the gas is determined by the difference between the enthalpy at the point where the equal pressure line of the second pressure crosses the saturated liquid line and the saturated gas line and the enthalpy at the point d (point s). Since the inverse ratio is obtained, the ratio between the weight flow rate of the liquid at the point e and the weight flow rate of the gas at the point h is equal to the line segment sh: line segment es. Further, as described below, since the liquid in the state at the point e reaches the point k via the evaporator 210,
The amount of gas at point k is equal to the amount of liquid at point e. Therefore, the position of the point m on the isobar is represented by the line segment h.
m: line segment mk = line segment sh: line segment es.

The refrigerant liquid at the point e separated by the gas-liquid separator 350 is supplied to the processing air cooler 30 which is the second heat exchanger.
0 evaporating section 251. Gas-liquid separator 35
0 and the evaporating sections 251A, B, C are connected by refrigerant pipes 430A, B, C, respectively. In this way, almost only the liquid flows into the evaporation sections 251A, B, C.

In the evaporating section 251, the refrigerant liquid evaporates under the second pressure and reaches a point f between the saturated liquid line and the saturated gas line at the same pressure. Here, most of the liquid may have evaporated, or some of the liquid may remain (FIG. 4).
(A) shows a state in which some liquid remains.)

The refrigerant in the state indicated by the point f flows into the condensing section 252. In the condensing section 252, the refrigerant is deprived of heat by the outside air and / or sprayed water flowing through the second compartment and reaches point g. This point is on the saturated liquid line in the Mollier diagram. Temperature 40 ° C, enthalpy 1
13.51 kcal / kg. The point e and the point g substantially match. The only difference is the pressure loss when the refrigerant flows through the evaporating section 251 and the condensing section 252.

The refrigerant liquid at the point g is supplied to the throttle 270 at a temperature of 10 ° C.
4.2 kg (as third pressure) which is the saturation pressure of ° C.
/ Cm ^2, and reaches a refrigerant evaporator 210 as a mixture of a refrigerant liquid and a gas at 10 ° C., where it takes heat from the processing air and evaporates to become a saturated gas in the state of point a on the Mollier diagram. Is sucked into the compressor 260 again, and the above cycle is repeated.

As described above, the processing air cooler 30
In 0, the refrigerant evaporates from the point e to the point f in the evaporating section 251, and changes state from the point f to the point g in the condensing section 252. Very high heat transfer coefficient.

Further, as a compression heat pump including the compressor 260, the refrigerant condenser (regeneration air heater) 220, the throttles 360 and 270, and the refrigerant evaporator 210, the gas-liquid separator 350 and the processing air cooler 300 are not provided. In this case, the refrigerant in the state at the point d in the refrigerant condenser 220 is returned to the refrigerant evaporator 210 through the throttle, so that the enthalpy difference that can be used in the refrigerant evaporator 210 is 148.83-122.97.
= 25.86 kcal / kg, whereas in the case of the heat pump HP1 used in the present embodiment provided with the gas-liquid separator 350 and the processing air cooler 300, 148.
83-113.51 = 35.32 kcal / kg, and the gas amount circulating to the compressor for the same cooling load is
As a result, the required power can be significantly reduced (about 15%). This is a so-called economizer cycle.
That is, since the gas-liquid separator 350 and the processing air cooler 300 are provided, the heat exchange between the processing air and the outside air can be performed with a high heat transfer coefficient, and the required power can be greatly reduced.

Here, if the compressor 260 is a multi-stage compressor (regardless of whether it is a centrifugal compressor or a reciprocating compressor), an intermediate port 260a is provided between the stages to make the above cycle possible. Can be realized. In addition, even with a single-stage compressor, the above-described cycle can be realized by taking the intermediate port 260a by a method described later.

As described above, in the embodiment of the present invention, since the gas-liquid separator 350 is provided, the gas-liquid separator 350 is guided to the heat exchange tubes (heat transfer tubes) constituting the evaporation sections 251A, B, and C of the heat exchanger 300. Almost no gas phase is contained in the refrigerant. Therefore, the evaporating section 251A,
The amounts of the refrigerant guided to B and C become uniform, so that the cooling of the processing air as the first fluid by the evaporation in the evaporation sections 251A, B and C becomes uniform, and the condensation section 2
The amount of refrigerant condensed in the heat transfer tubes 52A, B, and C is occupied by the refrigerant evaporated in 251A, B, and C in the evaporating section.
When a gas phase is contained, non-uniform heat transfer occurs, particularly in a condensing section containing a large amount of gas phase, but such a problem does not occur if only a liquid layer is used.

In this way, the heat pipe function of each heat transfer tube (the phase change of the refrigerant, especially the heat transfer function by evaporation and condensation)
Since the amount of heat transferred by the heat becomes uniform between the heat transfer tubes,
The uniform heat transfer can be performed in the entire heat exchanger 300, and the inconvenience that air as the first fluid and the second fluid passes without being involved in the heat transfer can be prevented. Therefore, in the dehumidifying air conditioner using the heat exchanger 300,
It is possible to improve the heat exchange efficiency between the processing air as the first fluid and the cooling medium (outside air) or the regeneration air as the second fluid, and to improve the operation reliability.

Next, referring to a partial extraction diagram of the Mollier diagram shown in FIG. 4B (only the point h and the vicinity of the compression stroke by the compressor 260 are extracted), a third throttle 262 is provided. The Mollier diagram in the case will be described. Considering the second pressure and the suction pressure of the intermediate port 260a of the compressor, the second pressure is mainly governed by the temperature of the processing air as the first fluid and the temperature of the outside air as the second fluid. At this time, if the pressure of the intermediate port 260a is higher than the pressure determined by the temperatures of the processing air and the outside air, the intermediate port 260
No flow of refrigerant gas to a occurs. Therefore, compressor 2
The intermediate port pressure determined by the structure 60 (the opening position of the intermediate port 260a) needs to be designed to be lower than the pressure determined by the temperatures of the processing air and the outside air. The pressure difference (h-h ') between the two thus set is compensated by the throttle 262 such as an orifice, a manual valve, and a control valve provided in the refrigerant passage 340.

In the unlikely event that the pressure at the intermediate port 260a
Check valve (not shown) from the gas-liquid separator 350 to the intermediate port 260a so that the backflow of the refrigerant gas from the intermediate port 260a side to the gas-liquid separator 350 side does not occur even if the pressure becomes higher. Is preferably provided.

In FIG. 4B, the refrigerant gas in the state at the point h is decompressed by the throttle 262 with equal enthalpy to the pressure of the intermediate port 260a, and reaches the point h '.
Is mixed with the gas in the state of the point k ′ compressed in the first half of the above, and becomes the gas in the state of the point m ′. Then, it is compressed and reaches point b. With such a design, even if the temperature of the processing air or the outside air fluctuates somewhat, the stable heat exchange cycle between the processing air and the outside air and the operation of the economizer cycle can be performed.

Next, with reference to FIG. 5, a description will be given of a dehumidifying air conditioner according to a fourth embodiment incorporating a heat pump HP2 according to a third embodiment. The configuration and operation are the same as those of the first and second embodiments except that water is used as a cooling fluid flowing to the second section of the processing air cooler 300b. In the figure, in a cooling tower 470 installed outdoors, cooling water cooled to about 32 ° C. in summer passes through a cooling water pipe 471 connected to the bottom of the cooling tower 470 and a cooling water pump 46.
The cooling water pipe 472 connected to the suction port of the processing air cooler 300b is sent to the second section 320 of the processing air cooler 300b.

The second section 32 of the processing air cooler 300b
In the case of 0, the cooling water flows outside the heat exchange tube perpendicular to the tube by removing the baffle plate provided to be perpendicular to the heat exchange tube. A cooling water pipe 473 is connected to a cooling water outlet of the second section 320, and is configured to return the cooling water whose temperature has increased in the processing air cooler 300 b to the cooling tower 470. Thus, the first of FIG.
In the second embodiment, the refrigerant is condensed in the condensing section by the outside air, whereas in the third and fourth embodiments, the refrigerant is condensed in the condensing section by the cooling water. The refrigerant cycle of the heat pump HP2 is shown in FIG.
The description is omitted because it is the same as that described above.

Next, referring to the table of FIG. 6, the second embodiment of the present invention will be described.
The operation mode of the dehumidifying air conditioner and the operation of each device according to the embodiment will be described. As shown in the table, the dehumidifying air conditioner of the second embodiment can operate in the cooling operation mode and the dehumidification operation mode. In the cooling operation mode, all of the desiccant rotor 103, the blower 102, the blower 140, the blower 160, the water spray 325, and the compressor 260 are operated or operated. The flows of the cooling fluid, the refrigerant, and the like are as described above.

In the dehumidifying mode, the desiccant rotor 10
3. The blower 102, the blower 140, and the compressor 260 are operating, but the blower 160 is stopped, and the water spray 325 is not operating. At this time, in FIG.
Since the outside air C, which is the cooling fluid, is not flowing, and water is not sprayed to the second section 320, little heat is taken from the refrigerant between the throttles 360 and 270. However, at this time, the second pressure is set at the intermediate port 260a of the compressor 260.
The processing air is cooled by the temperature corresponding to the suction pressure, and the refrigerant evaporated due to the cooling is also evaporated through the expansion valve 270 into the evaporator 2.
Since it is sent to 10, the processing air is not much different from the case where it is cooled only by the evaporator 210.

Therefore, considering the psychrometric chart of FIG. 3, there is almost no cooling between the state L and the state M,
Since the processing air is only cooled by the refrigerant evaporator 210 after being dehumidified by the desiccant rotor 103, the state in which the processing air is returned to the air-conditioned space has a lower absolute humidity than the state K, and the dry bulb temperature is lower. This is a state that is not much different from the state K. That is, this operation mode is basically a dehumidification operation mode. In the third and fourth embodiments of FIG. 5, if the cooling water pump 460 is stopped, the same dehumidifying operation mode as described above is possible. Also, the gas-liquid separator 350 operates in the dehumidifying operation mode, so that the economizer cycle is effective.

FIG. 7 shows an example of a typical mechanical arrangement of a dehumidifying air-conditioning apparatus according to the second embodiment including the heat pump HP1 according to the first embodiment having the flow shown in FIG. Show. In this mechanical arrangement, the heat exchanger 300
Is located above the desiccant rotor 103 in the vertical direction,
10 are arranged vertically below. Heater 220
Is disposed horizontally beside the cooler 210 and above the desiccant rotor 103 whose rotation axis is oriented vertically. The gas-liquid separator 350 is arranged below the heater 220 and near the first section 310 of the heat exchanger 300. The second section 320 includes the desiccant rotor 10
3 and is disposed in a processing air flow path that flows upward from below in the vertical direction.

In such a structure, the refrigerant condensed by the heater (refrigerant condenser) 220 flows into the gas-liquid separator 350 via the expansion valve 360, and the separated refrigerant liquid is supplied to the first refrigerant liquid.
Into the evaporation section in the compartment. Then it is condensed in the condensing section. The gas-phase refrigerant separated by the gas-liquid separator 350 is guided to the intermediate port 260a of the compressor 260, where it is sucked into the compressor 260.

Referring to FIG. 8, a specific example in which a scroll compressor is used as a preferred compressor used in the embodiment of the present invention will be described. This figure is a cross-sectional view as seen in the axial direction of the rotary shaft of the scroll compressor. The scroll compressor performs a fluid compression and a pump action by a combination of two spirals. In the figure, spiral scroll 2
60d is a fixed scroll relatively fixed to a casing (not shown). Another spiral oscillating scroll 26
0c is assembled with the fixed scroll 260d so that the centers of the two scrolls substantially coincide with each other. A discharge port 260b is provided in a casing (not shown) substantially at the center.

The swing scroll 260c swings in contact with the fixed scroll 260d. The contact position 2
A crescent-shaped compression chamber 26 between the location and the two scrolls
0e is formed. The compression chamber 260e moves from the peripheral portion toward the central portion, that is, toward the discharge port 260b while gradually reducing the volume by the swinging motion of the swinging scroll. As a result, the fluid is compressed and discharged from the outlet 260b. The crescent-shaped compression chambers are formed in pairs of the same shape and are united at the center. Oscillating scroll 26
The fluid is sucked in the peripheral portion by the movement of 0c, and the suction stroke is completed by forming a crescent-shaped compression chamber 260e. The number of turns of the scroll is usually around three, and in this case, three pairs of compression chambers are formed.
In the flow of the fluid, suction, compression, and discharge are continuously performed in one direction. The scroll shape is a combination of a circular involute and a circular arc. The greater the number of spirals used in the scroll compressor, the smaller the pulsation width per revolution.

In the figure, the intermediate port 260a indicated by the broken line
It is provided at an intermediate position between a peripheral portion and a central portion of a casing (not shown). This position may be determined according to the desired intermediate pressure. The closer to the center, the higher the intermediate pressure.

Although the scroll compressor described above is of the swing type, it may be of a rotary type in which both scrolls of the pair rotate around their respective centers to perform a compression action or a pump action. In this case, for example, the scroll 260c may be a driving scroll connected to a driving shaft, and the scroll 260d may be a driven scroll driven by the driving scroll 260c. At this time, the relative positional relationship between the two scrolls is the same as that in the case of the swing type, and thus the duplicate description will be omitted.

Referring to FIG. 9, a specific example in which a screw compressor is used as a preferred compressor used in the embodiment of the present invention will be described. This figure shows the screw compressor 2
FIG. 3 is a partial cross-sectional perspective view showing a meshing state of two screws. In the figure, a casing 261h accommodating the screw of the screw compressor 261 is shown with its front part cut off.

In the figure, a pair of rotors having a twisted tooth profile, a male rotor 261c having a convex tooth profile, and a female rotor 261d having a concave tooth profile are provided inside a casing 261h.
And is incorporated. Each rotor is casing 261h
Are supported in parallel by bearings at both ends. In the case of the oilless type, a synchronous gear is disposed at the shaft end so that the gap between the rotors can be maintained at an appropriate value and the rotors can rotate without contacting each other.

As indicated by the arrow in the figure, the male rotor 261c arranged on the upper side rotates counterclockwise when viewed from the left side in the figure, and the female rotor 261d arranged on the lower side rotates clockwise. It is configured as follows. At this time, the discharge end 261b moves rightward in the figure, and the suction end 2 moves leftward.
61 g are located. The suction port 261f is located at the left end of the casing.

In such a configuration, when the rotor is rotated in the above-described direction, the space formed by the male rotor 261c, the female rotor 261d, and the casing 261h becomes 1 'from left to right in the figure. -1 'to 2'-
2 ', 3'-3', 4'-4 ', increasing with rotation. As the space increases, gas is sucked into the space from the suction end side. In a state where the gas is sufficiently sucked, the space between the male and female rotors is shut off from the suction end and continues to rotate, and this time compresses the sucked gas.

As described above, the suction, compression, and discharge processes are performed successively by the tooth profile, so that torque fluctuation and pulsation of gas flow are small, and because the balance of the rotating body is good, vibration is low, and high-speed It is suitable for rotation and can be made compact.

In such a screw compressor, an intermediate port 261a (shown by a broken line) is provided in a casing 261h at an intermediate position between the discharge end 261b and the suction end 261g.
Is provided. The pressure of the intermediate port 261a is any pressure intermediate between the suction pressure and the discharge pressure.
The pressure in the intermediate port can be freely selected depending on the position in the axial direction of the rotor. Such a screw compressor 261 is used as the compressor 2 in FIG. 1 or FIG.
60 can be used instead.

[0093]

As described above, according to the present invention, since the second heat exchanger for evaporating and condensing the refrigerant under the second pressure lower than the first pressure is provided, high heat transfer is provided. Rate heat exchange is performed, and the refrigerant gas generated under the second pressure is separated and sucked into the intermediate port of the compressor.
Can stabilize the heat exchange of the heat exchanger, and can increase the enthalpy difference per unit amount of the refrigerant, thereby providing a heat pump with significantly improved COP.

If such a heat pump is combined with a moisture adsorbing device, the heat pump can be used to cool the first fluid to which the moisture has been adsorbed by the moisture adsorbing device, and the dehumidification with significantly improved efficiency can be achieved. A device can be provided.

[Brief description of the drawings]

FIG. 1 is a flowchart of a dehumidifying air conditioner according to a second embodiment, including a heat pump according to the first embodiment of the present invention.

FIG. 2 is a schematic sectional view of a heat exchanger suitable for use in the heat pump of FIG.

FIG. 3 is a psychrometric chart for explaining the operation of the dehumidifying air conditioner of FIG. 1;

FIG. 4 is a Mollier diagram of the heat pump according to the first embodiment.

FIG. 5 is a flowchart of a dehumidifying air conditioner according to a fourth embodiment including a heat pump according to a third embodiment of the present invention.

FIG. 6 is a table showing an operation mode of a dehumidifying air conditioner according to an embodiment of the present invention and an operation of each device.

FIG. 7 is a schematic front sectional view showing an example of an actual structure of a dehumidifying air conditioner according to a second embodiment of the present invention.

FIG. 8 is a schematic sectional view of a scroll compressor suitable for use in the embodiment of the present invention.

FIG. 9 is a schematic partial cross-sectional perspective view of a screw compressor suitable for use in the embodiment of the present invention.

FIG. 10 is a flowchart of a conventional dehumidifying air conditioner.

11 is a psychrometric chart explaining the operation of the conventional dehumidifying air conditioner shown in FIG.

FIG. 12 is a Mollier diagram of a heat pump used in the conventional dehumidifying air conditioner shown in FIG.

FIG. 13 is a perspective view showing an example of the structure of a desiccant rotor.

FIG. 14 is a perspective view showing an example of a cross-flow heat exchanger.

[Explanation of symbols]

 101 Air-conditioned space 102, 140, 160 Blower 103 Desiccant rotor 121 Heat exchanger 165 Vaporization humidifier 210 Refrigerant evaporator 220 Refrigerant condenser 251A, 251B, 251C Evaporation section 252A, 252B, 252C Condensing section 260, 261 Compressor 262 Restrictor 270 Restrictors 300, 300b Process air cooler 310 First section 320 Second section 325 Watering pipe 350 Gas-liquid separator 360 Restrictor 470 Cooling towers 501, 502, 503 Filter HP1, HP2 Heat pump

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) F25B 6/04 F25B 6/04 B F28D 21/00 F28D 21/00 Z

Claims (5)

[Claims] 1. A refrigerant compressor for compressing a refrigerant; a first heat exchanger for removing heat from the refrigerant compressed by the refrigerant compressor by a high-temperature fluid and condensing the refrigerant under a first pressure;
A first throttle for reducing the pressure of the refrigerant condensed in the first heat exchanger to a second pressure; a refrigerant reduced in the first throttle by heat from a first fluid under the second pressure; A second heat exchanger for removing heat from the refrigerant by the second fluid and condensing the refrigerant after evaporating; a third heat exchanger for condensing the refrigerant in the second heat exchanger. And a third heat exchanger configured to apply heat from the low-temperature fluid under the third pressure to evaporate the refrigerant depressurized by the second throttle. And a gas-liquid separator between the first throttle and the second heat exchanger for separating the refrigerant decompressed to the second pressure into a refrigerant liquid and a refrigerant gas; The refrigerant compressor has an intermediate port that sucks gas at a pressure intermediate between the first pressure and the third pressure, Characterized in that the refrigerant gas separated in the liquid separator configured to be inhaled into the intermediate port; heat pump. 2. The heat pump according to claim 1, further comprising a third throttle between the gas-liquid separator and the intermediate port. 3. The second heat exchanger, wherein: a first section through which the first fluid flows; a second section through which the second fluid flows; and wherein the second heat exchanger penetrates through the first section. A first fluid flow path through which the refrigerant exchanges heat with the first fluid; and a second fluid flow path through which the refrigerant exchanges heat with the second fluid passes through the second section. The refrigerant flows from the first fluid flow path through the second fluid flow path, and the refrigerant flows under the second pressure on the flow side heat transfer surface of the first fluid flow path; 3. The method according to claim 1, wherein the refrigerant evaporates at the heat transfer surface of the second fluid flow path and the refrigerant is condensed substantially under the second pressure. The heat pump according to 1. 4. The heat pump according to claim 1, wherein the heat pump is provided upstream of a flow of the first fluid with respect to the second heat exchanger, And a moisture adsorbing device having a desiccant for adsorbing moisture in the fluid. A first step of condensing the refrigerant at a first pressure; a second step of depressurizing the refrigerant condensed at the first pressure to a second pressure to evaporate a part of the refrigerant; A third step of separating a part of the refrigerant gas evaporated in the second step from the remaining refrigerant liquid; evaporating the separated refrigerant liquid substantially under the second pressure, and first evaporating the refrigerant liquid by the evaporating refrigerant. A fourth step of cooling the fluid; a fifth step of condensing the refrigerant gas evaporated in the fourth step substantially under the second pressure;
A sixth step of depressurizing the refrigerant condensed in the step to a third pressure; evaporating the refrigerant depressurized in the sixth step to the first pressure;
A step of cooling the refrigerant gas evaporated in the seventh step; an eighth step of increasing the pressure of the refrigerant gas evaporated in the seventh step to the first pressure; and a pressure increase of the refrigerant gas separated in the third step in the eighth step. A ninth step of merging with the refrigerant gas to be performed during the pressurization; a tenth step of adsorbing moisture in the first fluid with a desiccant; and heating the high-temperature fluid by the refrigerant condensed in the first step. An eleventh step of desorbing moisture from the desiccant adsorbed in the tenth step with the heated high-temperature fluid to regenerate the desiccant; a method of dehumidifying a fluid.