JP4617958B2 - Air conditioner - Google Patents

Description

  The present invention relates to an improvement in the operation efficiency of an apparatus that circulates refrigerant by a vapor compression refrigeration cycle and performs efficient air conditioning, for example, an air conditioner that can adjust indoor temperature and humidity.

  In this type of conventional air conditioner, that is, an air conditioner that controls both temperature and humidity, there are those that generate different evaporating temperatures in two independent refrigeration cycles having different capacities to control temperature and humidity (for example, , See Patent Document 1). Further, a technique is known in which temperature and humidity control is performed by cooling using a refrigeration cycle and a dehumidifier using the exhaust heat, and the operation efficiency of the air conditioning system is improved by linking and controlling them (for example, Patent Document 2). reference). Further, a technique is known in which the arrangement of a sensible heat cooling unit provided indoors can be freely performed by introducing outside air. (For example, see Patent Document 3)

JP-A-10-259944 (Claim 1, 0100 column, FIG. 1, FIG. 2, etc.) Japanese Patent Laid-Open No. 2002-22245 (column 0033, FIG. 1, etc.) Japanese Patent Laying-Open No. 2005-49059 (Claim 1, FIG. 2, etc.)

However, when a plurality of refrigeration cycles are installed in the same room, there is a problem that the cost of refrigerant piping construction increases. In addition, in a refrigeration cycle that operates at a high sensible heat ratio, that is, a refrigeration cycle that operates at a high evaporation temperature, an efficient operation is possible, but an operation at a low sensible heat ratio, that is, an operation at a low evaporation temperature. In the refrigeration cycle on the side, the compression ratio becomes large and the operation consumes a large amount of power.

In addition, when the refrigeration cycle operation control is performed so as to maximize the coefficient of performance (COP) in an air conditioning system equipped with a dehumidifier utilizing exhaust heat of the refrigeration cycle, the dehumidification unit is not limited to the refrigerant pipe to the indoor unit. However, there is a problem that the construction cost increases. There is also a problem that power consumption cannot be reduced even if each unit is optimally operated.

  The object of the present invention is to solve the above-mentioned problems, and the present invention can reduce the work load by using one refrigerant pipe connecting each unit as a system, so that a predetermined cooling load can be achieved. On the other hand, an object is to obtain an air conditioner capable of operating at maximum efficiency. Another object of the present invention is to obtain an efficient air conditioner having a simple structure that can improve environmental problems.

  In the air conditioner according to the present invention, a first compressor that discharges refrigerant, a heat source side heat exchanger, a first pressure reducing unit, and a first load side heat exchanger are sequentially connected to circulate the refrigerant. A refrigerant cycle, a second decompression means connected in parallel with the first decompression means and the first load-side heat exchanger, a second load-side heat exchanger and a first compression in the first refrigerant cycle The air conditioned by the second refrigerant cycle having the second compressor that is operated independently of the compressor, the first load side heat exchanger, and the second load side heat exchange in the same air conditioning region A load-side air blowing means for blowing out, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is dew point of the air-conditioning region The heat treatment capacity of the first load-side heat exchanger is adjusted to the heat treatment of the second load-side heat exchanger so as to be lower than the temperature. It is made larger than the force.

  The air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, a first pressure reducing unit, and a first load side heat exchanger, which are sequentially connected to circulate the refrigerant. A refrigerant cycle and a first decompression unit are formed, and a second decompression unit and a second load-side heat exchanger are connected in parallel with the first decompression unit, and the first decompression unit is driven during decompression. The pressure-increasing means for increasing the refrigerant pressure from the second load-side heat exchanger using suction by flow is the same as the air conditioned by the first load-side heat exchanger and the second load-side heat exchange Load-side air blowing means that blows out to the air-conditioning region, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is The heat treatment capacity of the first load-side heat exchanger is set so that it is lower than the dew point temperature of the air-conditioning area. It is made larger than the heat treatment capacity of the exchanger.

  An air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, a first The capacity can be adjusted independently of the first indoor unit including the indoor blower, the second decompression means, the second indoor heat exchanger, the second indoor blower with adjustable air volume, and the first compressor. A second indoor unit comprising a second compressor, an outdoor unit, a first indoor unit, and a refrigerant pipe connecting the second indoor unit, and a first indoor heat exchanger surface temperature Is operated so as to be higher than the dew point temperature of room air.

  An air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, a first A second indoor unit including a first indoor unit including an indoor fan, an expansion power recovery unit, a second decompression unit, a second indoor heat exchanger, a second indoor fan, and a second compressor. The unit, the outdoor unit, the first indoor unit, and the second indoor unit are connected by the high pressure side liquid pipe and the low pressure side gas pipe, and the second compression is performed by the power recovered by the expansion power recovery means by the high pressure side refrigerant. It drives the machine.

  The air conditioner according to the present invention is provided with a separate heat exchanger for processing sensible heat and latent heat to perform air conditioning, so that it is possible to operate efficiently with a simple structure. In the present invention, when the same cooling load is processed with two different evaporating pressures of the low stage and the high stage, the low stage side evaporating temperature is minimized so that the power consumption is minimized. Can do.

  Further, according to the present invention, since the air is pre-cooled by the high-stage heat exchanger and then cooled and dehumidified by the low-stage heat exchanger, the refrigerant flow ratio on the low-stage side is smaller than that on the high-stage side, and the total consumption is reduced. Electric power can be reduced. In addition, the present invention enables an efficient heat exchanger structure and arrangement according to the purpose and characteristics.

Embodiment 1 FIG.
FIG. 1 shows an example of a refrigerant circuit of an air conditioner according to an embodiment of the present invention. In FIG. 1, an outdoor unit 1 provided on the outdoor side or the like, which is on the heat source side, a first indoor unit 2 which is disposed on the indoor wall surface or the like and is on the load side. This is the second indoor unit 3. The first and second indoor units 2 and 3 are arranged so as to process the air conditioning load in the same space as the air conditioning area. Good. The encapsulated refrigerant is R410A, which is a fluorocarbon refrigerant, or a natural refrigerant such as carbon dioxide, isobutane, or water. FIG. 1 illustrates the configuration in which the refrigerant is circulated in the direction of cooling in the indoor unit, but there are four directions at the discharge ports of the compressors 4 and 14 provided in the outdoor unit 1 and the second indoor unit 3, respectively. It goes without saying that the refrigerant circuit can be changed to a refrigeration cycle that can be heated by providing a valve and switching the discharge direction to reverse the circulating refrigerant circuit. The first compressor is explained as being capable of adjusting the capacity by driving the inverter. However, the compressor is controlled by turning the compressor on and off using an induction motor and the like that do not adjust the rotational speed that always keeps operating at the maximum capacity. Alternatively, an efficient operation and a device having a simple structure can be obtained.

  The outdoor unit 1 includes a high-stage compressor 4 that sucks and discharges a refrigerant such as a scroll or a rotary that can adjust its operating capacity, an outdoor heat exchanger 5 that exchanges heat with outdoor air, and an outdoor fan 6 that adjusts the amount of heat exchange. Is done. Reference numerals 7 and 8 denote refrigerant pipes that connect the outdoor unit 1 and the indoor units 2 and 3. The high-temperature and high-pressure gas refrigerant discharged from the high-stage compressor 4 condenses in the outdoor heat exchanger 5, exchanges heat with the outdoor air blown by the outdoor blower 6, becomes high-pressure liquid refrigerant, and is carried indoors through the liquid pipe 7.

  The first indoor unit 2 includes a first pressure reducing device 9, a first indoor heat exchanger 10, and a first indoor blower 11. The first indoor unit 2 has a heat transfer area and an air blowing amount that are sufficiently large in heat exchange capacity with respect to the heat exchange amount, and the temperature of the refrigerant flowing through the first indoor heat exchanger 10 is the indoor air. Above dew point temperature. Therefore, in this 1st indoor unit 2, indoor air is not dehumidified but only a sensible heat load is processed. Here, the first indoor heat exchanger and the first indoor fan of the first indoor unit 2 have sufficiently large heat transfer performance (heat transfer area and air flow rate) with respect to the heat exchange amount. For example, heat exchange is performed between the refrigerant passing through the interior and the indoor air blown by the first indoor blower, but the heat transfer area and the air flow rate are set so that the temperature difference between the refrigerant and the indoor air is within 10 degrees. That is. For example, when the intake temperature of indoor air is 27 ° C., the humidity is 50%, and the dew point temperature is 15 ° C., the intake air temperature-evaporation temperature is about 10 ° C. or less, so that dehumidification is hardly performed during cooling and the evaporation temperature is high. Operation with a low compression ratio and high efficiency is possible. As described above, the heat transfer performance of the heat exchanger is determined by the heat transfer area and the air volume for heat exchange, and if the heat transfer performance is such that the difference between the air temperature and the evaporation temperature is within about 10 degrees, the evaporation temperature is almost Above dew point temperature. The refrigerant temperature is usually measured by a tube temperature thermometer provided on the heat transfer tube surface of the heat exchanger.

  Since the sensible heat exchanger used in the first indoor unit has a dew point temperature or higher, there is almost no condensation on the surface of the heat exchanger, and there is no need for a dew condensation water treatment mechanism such as a drain pan, drain pump, drain pipe, etc. Alternatively, it is possible to provide a simple structure in which only a drain pan having a shallow bottom is provided and the condensed water slightly accumulated is naturally evaporated. Therefore, the configuration of the first indoor unit and the surroundings of the first load side heat exchanger are simplified, and the equipment cost and the construction cost can be reduced. In this way, on the high stage side, the first compressor is controlled by the difference between the target set temperature of the room air and the actual room air temperature to be measured. For this reason, when the room is abnormally high in humidity, the sensible heat exchanger, which is the first load-side heat exchanger, has the possibility of causing condensation. A drain pan is provided, or a simple flow path that allows some condensed water to flow through the cooling fins even in the heat exchanger structure.

  Next, the second indoor unit 3 includes a second decompression device 12, a second indoor heat exchanger 13, a low-stage compressor 14, and a second indoor blower 15, and is connected to the liquid pipe 7 and the gas pipe 8 in the first room. It is connected in parallel with the unit, and the refrigerant temperature here is operated so as to be equal to or lower than the dew point temperature of the room air. Therefore, both the sensible heat load and the latent heat load are processed in the second indoor unit 3, and dehumidification is performed along with cooling of the indoor air during the cooling operation in which the outdoor unit is condensed. For this reason, a countermeasure against condensed water is necessary for the second load-side heat exchanger. As described above, in the refrigeration cycle of the present invention, when the outdoor unit 1 performs the condensation operation, the two indoor units perform the evaporation operation at different evaporation temperatures by changing the opening settings of the expanders 9 and 12. When used for an air conditioner, the higher evaporation temperature uses a heat exchanger with a larger area to operate mainly by sensible heat, and the lower evaporation temperature uses the air volume of the blowing means as in the latent heat main operation. The amount of heat treated can be controlled by the opening degree of the decompression means, etc., so that the cooling operation and the dehumidification operation are performed with high efficiency. Further, a four-way valve combined with each compressor can be provided to perform heating operation, that is, each indoor unit can perform condensation operation.

  Next, the refrigeration cycle operation of the air conditioner of the first embodiment configured as described above will be described with reference to FIG. 1 and FIG. FIG. 2 is a Ph diagram illustrating the refrigeration cycle operation by the air conditioner of the first embodiment. The horizontal axis represents specific enthalpy [kJ / kg], and the vertical axis represents refrigerant pressure [MPa]. Further, G1 and G2 indicated by arrows in FIG. 2 respectively indicate the refrigerant flow rate [kg / h].

  The high-temperature and high-pressure gas refrigerant (state A) discharged from the high-stage compressor 4 dissipates heat to the outside air by the outdoor fan 6 in the outdoor heat exchanger 5 and condenses to become high-pressure liquid refrigerant (state B). This high-pressure liquid refrigerant flows through the refrigerant pipe 7 and is decompressed by the first decompression device 9 and decompressed to an intermediate pressure Pe1 (state C) of about 20 ° C. at the saturation temperature. Exchange and evaporate. On the other hand, in the second pressure reducing device 12, the pressure is reduced to a low pressure Pe2 (state E) of about 5 ° C. at the saturation temperature, and the second indoor heat exchanger 13 exchanges heat with room air to evaporate. This low-pressure gas refrigerant (state F) is pressurized to the intermediate pressure Pe1 by the low-stage compressor 14 (state G), and merges with the medium-pressure gas refrigerant (state D) that has flowed out of the first indoor heat exchanger 10 (state D). H), and again sucked into the high stage compressor 4.

  Next, the coefficient of performance (COP) of the refrigeration cycle in the first embodiment shown in FIG. 2 will be described. The COP is obtained by dividing the cooling capacity of the refrigeration cycle by the compressor input, and the higher the value, the higher the efficiency of the refrigeration cycle.

As indicated by an arrow in FIG. 2, when the refrigerant flow rate flowing through the first indoor heat exchanger is G1, and the refrigerant flow rate flowing through the second indoor heat exchanger is G2, the low-stage compressor 14 uses the refrigerant flow rate G2. And the refrigerant flow rate (G1 + G2) after merging is increased in the high stage compressor 4. If Δh1 and Δh2 are compression power per unit refrigerant flow rate on the high stage side and the low stage side, the input of the low stage compressor 14 is G2 × Δh2, and the input of the high stage compressor 4 is (G1 + G2) × Δh1. Given.

The cooling capacity of this refrigeration cycle is obtained by the product of the evaporator enthalpy difference Δhe and the refrigerant flow rate. Here, if the refrigerant is HFC, the outlet enthalpies (state D and state F) of the first indoor heat exchanger and the second indoor heat exchanger are almost equal, so the evaporation enthalpy difference Δhe in each heat exchanger is equal. Then, the cooling capacity is (G1 + G2) × Δhe. Thus, COP is given by equation (1).
COP = (G1 + G2) Δhe / ((G1 + G2) Δh1 + G2Δh2) (1)

  This equation (1) indicates that when the cooling capacity (G1 + G2) Δhe of this air conditioner is constant, the input of the high-stage compressor 4 is determined only by Δh1 regardless of the ratio of G1 and G2, and the low-stage compression The machine input shows that the smaller G2 and the smaller Δh2, the smaller. Therefore, to increase the COP of this refrigeration cycle, it can be said that Δh1 is decreased, that is, Pe1 is increased, Δh2 is decreased, and G2 is decreased under a constant (G1 + G2). . However, since the high stage evaporation pressure Pe1 is determined by the heat transfer performance of the first indoor heat exchanger 10, it cannot be increased to a certain value or more.

  On the other hand, the dehumidification characteristic in the second indoor heat exchanger on the lower stage side is the sensible heat ratio (the ratio of the sensible heat capacity to the total cooling capacity) as the evaporation temperature (Pe2) that is the evaporation pressure of the refrigerant increases. ) Increases, the refrigerant flow rate G2 for obtaining the same dehumidification amount increases. FIG. 3 is a graph showing the sensible heat ratio characteristics with respect to the evaporation temperature change of the low stage heat exchanger. The horizontal axis represents the evaporation temperature and the vertical axis represents the sensible heat ratio. The characteristics are shown for three types of suction temperature 27 ° C., suction relative humidity 50%, 60%, and 70%. As shown in this figure, as the evaporation temperature gradually approaches the dew point temperature of the suction air, the sensible heat ratio Rises rapidly. Therefore, by increasing the evaporation pressure Pe2 up to a certain point, a COP improvement effect can be obtained by reducing Δh2, but when the evaporation temperature rises above a predetermined value, the low-stage refrigerant flow rate G2 increases rapidly to obtain the same dehumidification amount. COP deteriorates. FIG. 4 is a diagram showing a change in the low-stage compressor input with respect to the low-stage evaporation temperature, and shows the change in the refrigerant flow rate, the compression power per unit flow rate, and the compressor input under the constant dehumidification amount as the evaporation temperature on the horizontal axis.

  As shown in FIG. 4, the increase in the evaporation pressure Pe2 has both the effect of improving the COP in which Δh2 is reduced and the effect of deteriorating the COP in which the refrigerant flow rate G2 is increased. It can be seen that there exists an optimum evaporating temperature that minimizes COP, in other words, that maximizes COP. However, here, the first indoor heat exchanger 10 has sufficiently large heat transfer performance, and the high-stage refrigerant flow rate G1 is sufficiently larger than the low-stage refrigerant flow rate G2, and the low-stage refrigerant flow rate is low. It is assumed that the high-stage evaporation pressure Pe1 is hardly affected by changes in G2.

  As described above, in the air conditioner according to the embodiment of the present invention, the optimum evaporation temperature at which the low-stage compressor input is minimized is estimated, and operation control is performed so that the actual evaporation temperature approaches that value. Next, an example of control of the air conditioner in the embodiment of the present invention will be described with reference to the control flowchart of FIG. The actual evaporating temperature directly measures the temperature of the pipe through which the refrigerant of each indoor heat exchanger that is an evaporator flows, and the temperature measured outside this pipe is taken as the evaporating temperature.

  Although illustration is omitted, a temperature sensor and a humidity sensor are installed in the room, and an indoor air condition (temperature T and humidity RH) is detected at any time in step S1. Also, a target set temperature T * and set humidity RH * are set by a user interface such as a remote controller. Furthermore, high and low evaporating temperatures are detected as the current refrigeration cycle operating state. (S1a, S1b, S1c)

  In step S2, the sensible heat load and the latent heat load are estimated from the set temperature and humidity and the current temperature and humidity. It is assumed that the humidity deviation ΔRH indicates the magnitude of the latent heat load, and the temperature deviation ΔT indicates the magnitude of the sensible heat load. As described above, the first indoor heat exchanger 10 processes only the sensible heat load, and the second indoor heat exchanger 13 processes both the sensible heat load and the latent heat load. It is appropriate to control the capacity of the high stage compressor 4 according to the capacity control of the machine 14 and the magnitude of the sensible heat load.

  In step S3, the optimal evaporation temperature Te2 * of the second indoor heat exchanger 13 is calculated. This Te2 * is a value obtained from the current indoor air condition and the magnitude of the latent heat load, and is obtained from the relationship shown in FIG. FIG. 6 is a diagram showing the relationship between the low-stage evaporation temperature and the COP. This shows the relationship between the evaporation temperature and the COP when the room air is 27 ° C. and 50%. The refrigeration cycle shown in FIG. The relationship between the whole COP and the low stage evaporation temperature is shown. The air temperature that can generate such characteristics, that is, the relationship between the heat load, temperature, and humidity is grasped by an air diagram, and the evaporation temperature Te2 * at which the COP becomes maximum can be calculated. SHF indicates a sensible heat ratio, which is a ratio of the sensible heat load to the cooling load to be processed as a whole, and calculates the evaporation temperature Te2 * which is the maximum COP when the evaporation temperature for each SHF changes. To find the most efficient target evaporation temperature. The latent heat load may be used as a parameter instead of SHF. For example, SHF 0.85 indicates that a latent heat load of 15% is generated, and the absolute value of COP increases as SHF increases because the latent heat load decreases and the low stage refrigerant flow rate G2 decreases. It is because it is doing.

  In step S4, as described above, the rotational speed of the high-stage compressor and the low-stage compressor is controlled according to the load, and the evaporation temperature Te2 in the second indoor heat exchanger 13 is set to the optimum value Te2 *. The air flow rate of the second indoor blower 15 is adjusted. Although not shown, the first indoor heat exchanger 10 operates more efficiently as its evaporation pressure Pe1 is higher. Therefore, the first indoor fan is controlled to operate at approximately the maximum air flow. If the target is reached in step 4, for example, if there is no dehumidification load, the role of the second compressor is terminated, and the target evaporation temperature of the second refrigerant cycle becomes equal to or higher than the dew point temperature. That is, after reaching step 4, the heat treatment is maintained as it is. When ΔRH, ΔT, and the like become zero, the compressor rotational speed maintains the value at that time, and the rotational speed decreases or stops only when the difference becomes negative.

  In the above description, in order to obtain the heat treatment capacity of the second load side heat exchanger that mainly processes the latent heat, the current humidity RH is measured with a hygrometer and the second compressor according to the deviation from the target humidity. However, the heat treatment capability may be obtained by a method in which the humidity is not directly measured. For example, the refrigerant temperature of the second load-side heat exchanger is lower than the dew point temperature from the dew point temperature obtained from the air diagram when the room humidity during cooling is a constant predetermined value and is estimated from only the room temperature. The air temperature for the heat exchanger or the second compressor is operated, or the tube temperature, which is the refrigerant temperature of the second load-side heat exchanger, is compared with the measured value of the air temperature so that the tube temperature is air. By controlling the blower and the refrigerant cycle so as to be approximately 10 degrees or more lower than the temperature, the latent heat treatment can be performed without having a humidity sensor.

  The variable expansion valve 9 serving as the first pressure reducing means and the variable expansion valve 12 serving as the second pressure reducing means have an outlet refrigerant superheat degree of the indoor heat exchangers 10 and 13 to which the respective one ends are connected are approximately 5 deg. The refrigerant flow rate is adjusted so that Note that the degree of supercooling may be used instead of the degree of superheat, and other than the compressor, such as an ejector for increasing the pressure, can be controlled by opening and closing with a needle valve.

  As described above, the air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, an outdoor heat exchanger, an outdoor unit including an outdoor fan, a first decompression unit, and a first indoor heat. Exchanger, first indoor unit including second indoor blower, second decompression means, second indoor heat exchanger, second indoor blower with adjustable air volume, second compression with adjustable capacity A second indoor unit provided with an air conditioner, a refrigerant pipe connecting them, an indoor temperature / humidity information detecting means, a refrigerant temperature detecting means flowing through the second indoor heat exchanger, and a target refrigerant temperature setting means. And the air flow rate of the second indoor blower is controlled so that the refrigerant temperature becomes the target refrigerant temperature, so that the power consumption for processing a predetermined cooling load can be minimized. Control devices such as a compressor rotation speed control device, an expansion valve opening control device, a target refrigerant temperature setting means, etc. are provided in necessary places such as the outdoor unit side or the indoor unit side, and are individually controlled, What is necessary is just to control collectively by a communication means in one place.

  FIG. 7 is a refrigerant circuit diagram and shows another refrigerant circuit of the present invention for reducing the low-stage refrigerant flow rate G2. A difference from the refrigerant circuit shown in FIG. 1 is that a third decompression means 16 and a third indoor heat exchanger 17 are arranged in the second indoor unit 3. As shown in FIG. 7, the third indoor heat exchanger 17 is disposed at a position where the same operation as that of the first indoor heat exchanger 10 is performed.

  By cooling and dehumidifying the air that has passed through the third indoor heat exchanger 17 with the second indoor heat exchanger 13 on the lower stage side, a part of the cooling heat amount is shifted to the higher stage side as shown in FIG. be able to. FIG. 8 shows the change in air condition from suction to blowing on the air diagram. A broken line is an air state change according to the configuration of FIG. 1 and changes from the state A to the state B, and the heat amount change per unit air mass is indicated by Δi1. 7 changes from the state A through the third indoor heat exchanger 17 to the state C, and further cooled by the second indoor heat exchanger 13 to the state B. An apparatus with high efficiency can be obtained by first processing the amount of heat during cooling by the blower 15 in the heat exchanger 17 and lowering the amount of sensible heat load to be processed in the heat exchanger 13.

  The state of the inlet air and the outlet air of the indoor unit 3 in the refrigeration cycle of FIGS. 1 and 7 of the present invention is substantially equal. That is, although the dehumidification amount ΔX is equal, it can be seen that the amount of heat processed by the low stage evaporation pressure is smaller by Δi2 in FIG. As explained in FIG. 2, the refrigeration cycle COP of the present embodiment under a certain cooling load is improved by reducing the ratio of the low-stage refrigerant flow rate G2 to the total refrigerant flow rate (G1 + G2). In this configuration, the refrigeration cycle COP of the present embodiment is further improved by shifting the required cooling capacity to the higher stage side.

  In the case of air conditioning for buildings, the law stipulates that outside air is forcibly taken in by machines. In such a case, the dehumidifying load is generated by introducing outside air having higher humidity than room air into the room. When a dehumidifying load is generated by forced ventilation, the second indoor unit directly processes the outside air, that is, by introducing the outside air into the second indoor unit, the ventilation of the second indoor heat exchanger 13 is performed simultaneously with the ventilation. The sensible heat ratio can be lowered, and the low-stage refrigerant flow rate G2 necessary for obtaining the same dehumidification amount can be reduced. That is, the lower stage refrigerant flow rate G2 required to obtain the same dehumidifying amount can be reduced by dehumidifying the high humidity outside air. FIG. 11 is a diagram for explaining the difference between when the outside air is cooled and dehumidified and when the room air is cooled and dehumidified. In FIG. 11, the state A is an indoor air state, and the process of cooling and dehumidification is indicated by solid arrows. In this case, in order to obtain a dehumidification amount of ΔX, a cooling capacity of Δi1 is necessary. On the other hand, the state B is outside air having a higher temperature and humidity than room air, and the process of cooling and dehumidification is indicated by broken-line arrows. In this case, the cooling capacity necessary to obtain a dehumidification amount of ΔX is Δi2. As shown in FIG. 11, in Δi1 and Δi2, Δi2 for dehumidifying high-humidity air is smaller. That is, the low stage refrigerant flow rate G2 for obtaining the same dehumidifying amount when the outside air having a higher humidity than the room is dehumidified may be small. When the outside air introduction load due to forced ventilation is inevitably generated, it is included in the cooling load to be processed. Therefore, an efficient apparatus can be obtained by directly treating the outside air with the latent heat exchanger and introducing it into the room.

  As a configuration for introducing the outside air into the room, the second indoor unit 3 shown in FIGS. 1 and 7 is arranged in the room as it is, and a structure in which an inlet for air sucked into the second indoor unit 3 is provided outside the room, There exists a structure which provides the blower outlet which blows off the air-conditioned air of a unit indoors as an external air conditioner arrange | positioned outdoors. The structure in which the inlet for introducing the outside air is provided outside the room requires a small hole in the wall, but the structure of the external air conditioner is simple from the construction work of the piping. The selection also depends on the indoor air conditioning performance such as the relationship with the position of the heat source in the room.

  FIG. 12 is a diagram for explaining the configuration of pre-cooling / reheating for the latent heat exchanger. As shown in FIG. 12, the lower stage side and the second load side heat exchanger 13 are arranged on the suction side. If the air is precooled in advance with another sensible heat exchanger or the total heat exchanger 56 and reheated with the air blown out from the latent heat exchanger 13, the efficiency of the apparatus is good. As shown in FIG. 12, a total heat heat exchanger 56 is provided, and after the indoor air is pre-cooled through the total heat heat exchanger 56 by a wind path structure as shown in the figure, that is, a blower and a wind path not shown, the latent heat heat exchanger 13 An air path is employed in which the low-temperature air that has been cooled and dehumidified is reheated through another air path of the total heat exchanger 56. At this time, the air blown out from the latent heat exchanger 13 is heated again and supplied to the room, but this temperature rise is cooled by the first indoor heat exchanger 10 that processes the sensible heat. The amount of heat treated by the exchanger 13 is reduced, and the cycle COP is improved. In FIG. 12, for example, indoor air at 27 ° C. is pre-cooled to 22 ° C. by the total heat exchanger 56 and then cooled and dehumidified to 12 ° C. by a latent heat exchanger which is a low-stage evaporator. Precooling from 27 ° C. to 22 ° C. is performed with air at 12 ° C. after this cooling and dehumidification. On the other hand, the air at 12 ° C. is reheated by the intake air at 27 ° C. and blown into the room at 17 ° C. Thereby, the efficiency of the latent heat treatment at the time of forced outside air intake can be improved.

  Instead of the total heat heat exchanger, a sensible heat exchanger such as a heat pipe may be provided on the suction side and the outlet side of the second load heat exchanger 13. FIG. 13 shows this state with an air diagram. The state A that is the intake air is pre-cooled to state C and is cooled and dehumidified. The air state B after dehumidification is reheated to become the state D and blows out again into the room. If there is no precooling and reheating, a cooling capacity corresponding to Δi1 is required. However, by performing precooling and reheating, the SHF of the heat exchanger for dehumidification becomes low, and the cooling capacity for obtaining the same dehumidification amount becomes Δi3, which is cooled by ΔI2. The ability becomes unnecessary.

  As described above, when the cooling load is processed with two pressures of the high-stage evaporation pressure and the low-stage evaporation pressure, the COP is improved by reducing the low-stage side refrigerant flow ratio as much as possible, but the dehumidification occupies the total cooling load. When the load ratio is approximately 0.4 or more, most of the total cooling load must be processed at a low stage evaporation pressure. As can be seen from the characteristics of the sensible heat ratio shown in FIG. 3, the sensible heat ratio does not become zero even if the evaporation temperature is further decreased, and the normal cooling setting conditions of 25 ° C. to 28 ° C., relative humidity of 40% to In the range of 60%, about 0.6 is the lower limit. The effect of the present invention is small in such a cooling load having a relatively high dehumidifying load ratio, and it can be said that the air conditioning system is suitable for an air-conditioning target having a dehumidifying load ratio of 0.2 or less with respect to the total cooling load. For example, the total cooling load of the first indoor heat exchanger and the second indoor heat exchanger is set to 100, and the latent heat load thereof is set to 40. Here, when the sensible heat ratio of the second indoor heat exchanger is 0.6, the load to be processed by the second heat exchanger is latent heat load 40 / (1-SHF) = 100, and the first indoor heat exchange is performed. The cooling sensible heat load to be processed by the cooler is lost, and the entire cooling load is processed by the second indoor heat exchanger, so that the energy saving effect cannot be obtained.

  Under normal cooling conditions, such as room temperature 25-28 ° C. and humidity of about 50%, sensible heat treatment can be performed with a temperature difference equal to or higher than the dew point temperature. On the other hand, below the dew point temperature, there is a lower limit on the state change line which is the state B in the direction inclined from the normal cooling condition to the refrigerant temperature direction of the saturated vapor pressure line as shown in the air diagram of FIG. There is no less than 0.6. In other words, it is impossible to process a linear heat load of 40% or more. Thus, it is desirable to design each heat exchanger area ratio so that the latent heat and sensible heat load treatment is not 40:60, that is, 2: 3, but a smaller latent heat treatment. Assuming that the load condition for obtaining the energy saving effect of the air conditioner is 10 to 20% of the latent heat load ratio with respect to the total cooling load, and the sensible heat ratio of the second indoor heat exchanger can be reduced to 0.5, The cooling load to be processed by the second indoor heat exchanger is 20-30% of the total cooling load. That is, the heat treatment capacity of the second heat exchanger relative to the first indoor heat exchanger is preferably about 20-30%. In other words, the heat exchanger on the sensible heat treatment side is more than three times the heat exchanger on the latent heat treatment side. That is. As described above, when the heat treatment capacity of the sensible heat exchanger is made larger than the heat treatment capacity of the latent heat exchanger, the heat transfer area ratio of the heat exchanger itself is increased by three times or the load side blowing means is generated. It is good also considering heat processing capability as 3 times or more by both air volume, ie, the rotation speed of a ventilation fan. Furthermore, if the maximum capacities of the first compressor and the second compressor are set to be three times or more, the entire system can be adapted during operation without causing a problem. For this reason, it is preferable to reduce the sensible heat ratio of the second indoor heat exchanger due to the small ratio of the cooling capacity to be processed by the second indoor heat exchanger and the air path configuration of the processing air as shown in FIG.

  Although the example which uses hydrofluorocarbon, for example, R410A as a refrigerant | coolant was shown in the above description, it is not limited to this. For example, in the case where a natural refrigerant such as carbon dioxide is used, the high-stage evaporation pressure is close to the critical pressure, and therefore the high-stage and low-stage evaporation enthalpy differences Δhe described with reference to FIG. No more. Next, a case where there is a difference in enthalpy difference that can be used in the evaporator at different evaporation pressures when the evaporation temperatures are different will be described.

  FIG. 9 shows a Ph diagram when carbon dioxide is used as the refrigerant of the air conditioner of the present embodiment. As shown in the figure, near the critical pressure, the saturated gas line is bent largely to the left on the high pressure side, so that the enthalpy difference that can be used on the high stage side is smaller than that on the low stage side (Δhe1 <Δhe2). This indicates that the G2 refrigerant flow rate ratio in the total refrigerant flow rate G1 + G2 is smaller than the low-stage processing heat amount ratio with respect to the total cooling load. That is, when 20% of the total cooling capacity is required to obtain the necessary dehumidification amount, Δhe is almost equal between the high stage and the low stage in terms of the physical properties of the HFC refrigerant as shown in FIG. (G1 + G2) also needs 20%, but according to the physical properties of the carbon dioxide refrigerant shown in FIG. 9, since Δhe2> Δhe1, G2 ratio may be smaller than 20%. This indicates that the carbon dioxide refrigerant has a higher COP improvement rate than the HFC refrigerant.

  Furthermore, as described above, the low-stage refrigerant flow rate G2 is 20% or less of the total refrigerant flow rate, which is a suitable condition for the air conditioner of this embodiment. By collecting the expansion power from high pressure to medium pressure, the low-stage compression power can be made zero. This will be described in detail with reference to FIG. Moreover, the same number is attached | subjected about the part similar to FIG. 2, and description is abbreviate | omitted. In particular, in the case of a refrigerant in a supercritical state, such as carbon dioxide, recovery of expansion power is effective because the pressure is high and there is a pressure difference.

The second indoor unit 3 incorporates expansion power recovery means 21. The expansion power recovery means 21 includes an expansion portion 22 and a compression portion 23, and is connected coaxially. The expansion power when the high-pressure liquid refrigerant flowing from the outdoor unit 1 is reduced to an intermediate pressure is recovered, and the compression unit 23 performs the compression work. In the example of FIG. 10, the refrigerant depressurized to the medium pressure is further depressurized by the second depressurization means 12, the indoor air is cooled and dehumidified by the second indoor heat exchanger 13, and then the pressure is increased again to the medium pressure by the compression unit 23. ing.

  Since the refrigerant flow rate compressed here is about 20% to 30% of the total refrigerant flow rate as described above, the flow rate can be sufficiently increased in view of the physical properties of carbon dioxide and the expansion power recovery efficiency. In this way, the low stage side has only expansion power, and other inputs can be made zero.

  The refrigeration cycle COP improvement method in these configurations is not limited to air conditioning. Even in a system that generates different evaporation temperatures by freezing and refrigeration, the COP can be improved by the same principle.

  The heat load is divided into sensible heat treatment and latent heat treatment as in the present invention, and the overall control of the blower, the arrangement of each processing unit, and the heat exchanger of the latent heat treatment are significantly larger than the heat exchanger of the sensible heat treatment. The efficiency improvement can be remarkably expanded by improving the overall efficiency by combining the structures such as reducing the size. As a result, a refrigeration cycle effective as a global environmental protection measure is obtained, which is effective for air conditioners, refrigerators, refrigeration warehouses, refrigeration systems in convenience stores and superstores, and the like. In addition, in a refrigeration cycle that uses a natural refrigerant such as carbon dioxide (CO2), the pressure is high, and as shown in FIG. Although consideration is necessary, the ozone layer is not destroyed like chlorofluorocarbons such as R22 on the global environment, and it has little impact on global warming like alternative chlorofluorocarbons. Therefore, the efficiency improvement is required with the carbon dioxide refrigerant, and it is desirable to use a large pressure difference based on the high pressure as the expansion power as shown in FIG. Next, another high efficiency measure in the refrigeration cycle using such a natural refrigerant will be described.

  As a mechanism for improving efficiency by using a large pressure difference, there is an expansion power recovery means 21, for example, an expansion portion 22 and a compression portion 23, and there is a mechanical power recovery structure that is connected coaxially and rotates at the same rotational speed. Such an expansion power recovery means 21 is an insulation of the enthalpy difference at the entrance of the expander, which is the expansion work associated with the expansion power volume expansion of the fluid that has been wasted as heat in the form of friction loss and vortex loss in the conventional throttle device. The heat drop is recovered and converted into mechanical energy such as rotational power or reciprocating power and used for the compression work of the compressor. The compression part 23 driven by the expansion part 22 of the expansion power recovery means 21 is used for indoor air conditioning as a second compressor. During cooling, the refrigerant is circulated to the second load-side heat exchanger that performs latent heat treatment, and during heating, power is recovered by the expansion unit 22 switched to the high-pressure side piping, and the refrigerant is circulated to the low-pressure side piping. A load process for blowing the heat of condensation of the second refrigerant cycle into the room is performed to help improve efficiency. In such a supercritical refrigerant at high pressure, the volume expansion on the high pressure side increases, that is, the density loss increases and the degree of thirst at low pressure increases, so the expansion loss is large, so the use of expansion power is effective for energy saving. It is. In addition, energy can be further saved by using only this expansion power for the expansion section 22 which is the driving force of the second compressor and returning the refrigerant to the high pressure side again. As described above, any structure can be used as long as it is a drive mechanism that uses the kinetic energy obtained from the pressure difference or density, which is the physical property of the refrigerant, that is, the fluid deposition flow rate or flow velocity.

  As described above, in the configuration of FIG. 10 according to the present invention, expansion work is performed when a high-pressure fluid expands to low pressure in the expansion section 22, and the work conventionally consumed as heat is driven by the driving force of the compression section 23. It is said. In particular, a refrigerant such as carbon dioxide gas that becomes a supercritical state at a high pressure can provide a relatively large recovery power in terms of physical properties. In addition, as a natural refrigerant for performing such work, there is a difference in enthalpy difference that can be used for an evaporator at different evaporating pressures even in a mixed refrigerant of natural refrigerant such as air, nitrogen, helium, or carbon dioxide gas and HFC or other refrigerants. Needless to say, a refrigerant is also effective. The expansion power recovery means 21 as an example is a simple structure in which the expansion portion 22 and the compression portion 23 are directly connected to the inside of the sealed container so as to directly compress the expansion power as it is. It is. That is, the expansion fixed scroll and the second compression fixed scroll are fixedly arranged opposite to each other in the axial direction in the sealed container, and a bearing for supporting the rotating shaft is provided at the center of the container. In the center of the container, a swing scroll for expansion and compression that meshes with each fixed scroll and rotates is driven by a rotary shaft and formed as an integral rotor. That is, it is a structure that is driven by the expansion power in the expansion portion and swings and rotates and performs compression work in the compression portion. For this reason, high-pressure refrigerant is sucked from the expansion suction pipe connected to the high-pressure pipe through the fixed expansion scroll through the fixed fixed scroll and into the position near the axial center of the meshing space with the expansion swing scroll, and the swing scroll is shaken. The rotor is rotated while being moved, and is discharged from the space close to the container wall to the expansion discharge pipe and returned to the high-pressure pipe again. Next, after passing through the expander 12, the refrigerant whose pressure has been reduced by evaporation in the second load side heat exchanger passes through the second compression suction pipe to the compression unit 23 through the compression fixed scroll and the compression swing scroll. As the gas refrigerant is sucked into the outer peripheral side deviating from the shaft center of the meshing space and the scroll of the compression portion 23 is driven and oscillated via the rotating shaft, the gas pressure is generated from the meshing space inside the shaft center. It is raised and returned from the discharge pipe to the low pressure pipe. That is, the expansion power can be converted into compression work by simply putting the refrigerant in and out of the respective scroll structures in the sealed container, and the operation without waste can be performed with a simple structure. Although the scroll compressor has been described here, a reciprocating compressor may be used. Further, if the impeller-shaped drive device using the volumetric flow rate and flow velocity of the fluid is used instead of the positive displacement type drive device, a bypass valve or the like for preventing the surplus which becomes a problem with the positive displacement type becomes unnecessary.

  Instead of such mechanical expansion power recovery means, it is possible to use an ejector for recovering power using pressure difference. FIG. 14 is a diagram for explaining the principle of this ejector. As shown in FIG. 14, when the high pressure refrigerant is decompressed and ejected from the nozzle 31, the low pressure refrigerant is sucked and mixed from the periphery of the nozzle to obtain a medium pressure refrigerant. The driving flow shown in the figure is a condensed high-pressure liquid refrigerant. When the driving flow is reduced in pressure and becomes a high-speed two-phase jet, the pressure becomes the lowest when it is ejected from the nozzle 31, and the suction flow is drawn. The combined driving flow and suction flow are mixed uniformly in the mixing section 32, decelerated in the diffuser section 33, recovered in pressure, and then proceed downstream. The ejector 30 loses energy due to the friction of the refrigerant flow with a normal pressure reducing means, for example, an expansion valve, etc., whereas this energy is used for boosting the suction flow. Therefore, when the drive flow is decompressed, the enthalpy As a result, the enthalpy that can be used in the evaporator can be expanded and a refrigerant flow lower than the compressor suction pressure can be generated. Further, by providing a needle valve at the nozzle 31 portion of the ejector, the refrigerant flow rate to the third indoor heat exchanger 17 can be controlled. That is, the ejector simply serves as an expansion valve of the sensible heat exchanger 17 by providing the needle valve.

  FIG. 15 shows a configuration diagram during cooling of a refrigeration cycle using the ejector 30. In the refrigeration cycle shown in the figure, the high-pressure refrigerant discharged from the compressor 14 and condensed in the outdoor heat exchanger 5 that is a condenser is sent out from the outdoor unit 1, and a part of the high-pressure refrigerant is ejected from the ejector 30 in the second indoor unit 3. The pressure is reduced in the section 31 and evaporated in the third indoor heat exchanger 17 as an evaporator, and mainly sensible heat of the indoor heat load is processed. The remainder of the high-pressure refrigerant is depressurized by the third depressurizing means 12, evaporated in the second indoor heat exchanger 13, and mainly processed for latent heat. The low-temperature and low-pressure refrigerant discharged from the second indoor heat exchanger is sucked into the high-pressure refrigerant from the vicinity of the nozzle 31 of the ejector 30 and becomes intermediate-pressure refrigerant, which is evaporated in the third indoor heat exchanger 17 and sucked into the compressor 4. Thus, the ejector 30 is operated so that the evaporation pressure of the second indoor heat exchanger 13 to which the suction part is connected is lower than the suction pressure of the compressor 4. As a result, even when the evaporation temperature in the third heat exchanger 17 is equal to or higher than the dew point temperature, the evaporation temperature of the second indoor heat exchanger becomes equal to or lower than the dew point temperature of the room air, and dehumidification is possible.

  Next, FIG. 16 shows a Ph diagram of a refrigeration cycle using an ejector, and FIG. 17 shows a refrigeration cycle configuration diagram for explaining this operation. FIG. 16 shows a simulated operation. A part of the high-pressure liquid refrigerant condensed in the condenser (state A) is decompressed to the state B by the second decompression means 12 and is subjected to the second indoor heat exchange. In room 13, the room air is evaporated while being cooled and dehumidified, so that state C is obtained. The refrigerant of the low-temperature gas in the state C is sucked into the ejector 30 and drawn into the mixing unit 32. On the other hand, the high-pressure liquid refrigerant in the state A flows into the drive flow of the ejector 30 and is depressurized while reducing the enthalpy by the pressurization work for sucking the suction flow, that is, the pressure increase from the state C to the state D in FIG. In the ejector, the refrigerants in the state D and the state E are merged to be in the state F, flow into the third indoor heat exchanger 17 and evaporate while cooling the indoor air to be in the state G. However, this operation assumes the state D and the state E for convenience of explanation, and actually moves characteristically from the state C to F and from the state A to F. In the ejector, the drive flow is reduced in pressure while the high-speed two-phase jet is ejected from the nozzle 31, the pressure is lowest and the suction flow is drawn. Therefore, the pressure outside the ejector evaporates in the pressurized state F, and the gas pipe The air is sucked into the first compressor 4 through 8.

  In FIG. 17, the third indoor heat exchanger 17 that mainly processes sensible heat in the second indoor blower 15 and the second indoor heat exchanger 13 that mainly processes latent heat exchange heat with room air by serial air blowing. Although the structure is described, this is to improve the efficiency characteristics of the latent heat exchanger, and the air to the indoor heat exchanger may be blown in parallel by one blower in consideration of the heat exchange area. In addition, an independent blower may be provided for each individual heat exchanger, and independent control may be performed depending on the operation status. FIG. 18 shows a configuration in which a latent heat exchanger and a sensible heat exchanger are accommodated in one indoor unit with a configuration different from that in FIG. 17, and a Ph diagram thereof is shown in FIG. The high-pressure liquid refrigerant discharged from the main compressor 4 that is the first compressor and condensed in the condenser 5 that is the heat source side heat exchanger provided outside the chamber, the state A is first reduced to the state B by the ejector. The temperature in state B is equal to or higher than the indoor dew point temperature. A part of the refrigerant in this state B is branched and further decompressed by the expansion valve 71 to a state C that is not more than the indoor dew point temperature. The refrigerant that has evaporated in the second indoor heat exchanger 13 and mainly processed the latent heat is in the state D and is pressurized by the ejector, and again reaches the temperature of the state B that is equal to or higher than the dew point temperature. The refrigerant in the state B is evaporated in the sensible heat treatment main body in the third indoor heat exchanger 17 to become the state E and is sucked into the compressor 4.

  In the configuration of FIG. 19, since the refrigerant on the inlet side of the expansion valve 71, that is, the refrigerant in the state B is two-phase, if vibration or noise due to the refrigerant flow when passing through the expansion valve becomes a problem, a practical structure is used. A configuration as shown in FIG. 20 is preferable. FIG. 20 is a diagram for explaining a part of the refrigeration cycle of FIG. 19, in which a distributor 72 is provided on the downstream side of the ejector, and a capillary tube having a large throttle amount, that is, a narrower tube, is provided on the second indoor heat exchanger side. After the capillary is connected and the pressure is reduced to the room dew point temperature or lower, the latent heat load is mainly processed in the second indoor heat exchanger to obtain the suction flow of the ejector 30. The connection to the third indoor heat exchanger 17 is a thick capillary tube with a small throttle amount, which ensures a refrigerant temperature equal to or higher than the indoor dew point temperature and mainly performs sensible heat processing. With this configuration, it is possible to form a refrigerant cycle system that can improve efficiency by using an ejector with a simple structure.

  18 and 20, the structure in which the blower 15 that flows the indoor air in series between the second indoor heat exchanger 13 and the third indoor heat exchanger 17 housed in the indoor unit 34 is described. As described above, each of the heat exchangers may be provided with a blower independently, or may be blown in parallel with a single blower. FIG. 21 shows the structure of the latter indoor unit. The air conditioner of FIG. 21 has a structure embedded in the ceiling 50 of the room. A blower 52 for sucking indoor air from the center is housed in the ceiling-embedded cassette 51, and the air blown around is cooled by a heat exchanger. Then it is blown out into the room again. The heat exchangers 10 and 13 are provided so as to cover the periphery of the blower 52 and collect condensed water by a drain pan 54 provided in the lower part. A filter 53 is provided in the intake air intake on the front side facing the room, and the back of the cassette 51, that is, an opening / closing valve that opens and closes the ejector suction path and bypass circuit during heating in a position hidden behind the ceiling, etc. A storage box is provided, which is housed and fixed and covered with a heat insulating and heat insulating material. The upper heat exchanger 10 (17) mainly processes sensible heat. It occupies most of the heat exchange area. Since the temperature of the refrigerant flowing through the heat exchanger is equal to or higher than the dew point temperature and there is almost no dew, there is no need to consider flowing water through the fins. Therefore, it is not necessary to arrange the fins side by side. The most efficient heat exchanger structure may be adopted. For example, a plurality of flat tubes are installed in the air flow direction that is squeezed up and down and spread horizontally as a heat transfer tube through which the refrigerant passes, and between the upper and lower heat transfer tubes Fins may be provided to improve efficiency. That is, the corrugated fin is sandwiched between the flat tubes to transfer heat. On the other hand, the lower heat exchanger 13 employs a structure in which the temperature of the refrigerant is set to the dew point temperature or less in order to process latent heat, and plate fins that are easy to treat condensed water are arranged in the vertical direction. Provided.

  Moreover, since the lower heat exchanger 13 mainly performs a latent heat treatment, it may have a multi-divided flow structure that increases the number of diverted heat transfer tubes. In order to improve efficiency in heat exchangers, the smaller the pipe diameter, the better. Therefore, efficiency can be improved even in a small area, and the number of diverts will increase. However, there is little fear of increasing unbalance. For such a thin tube heat exchanger, it is preferable that a wire-like fin is wound around or brought into contact with the heat transfer thin tube, and the wire is stretched obliquely so that water flows. In this way, the heat exchanger for sensible heat treatment and the heat exchanger for latent heat treatment are not only greatly different in size, which is the area of the heat exchange treatment, but also the size and shape of the heat transfer tubes for the purpose of improving their characteristics and efficiency, A combination of completely different fin dimensions and shapes can be used, and no special consideration is required for the combination connecting portion, and the fins can be simply arranged vertically as shown in FIG.

  As described above, the configuration of the refrigerant cycle in FIG. 1 and the like has been described mainly with respect to cooling. However, it is easy to heat indoor air conditioning by providing a four-way valve or the like that changes the flow of the refrigerant. An example of this configuration is shown in FIG. In FIG. 22, the high-stage compressor 4 and the outdoor heat exchanger 5, which are heat source side cycles, are arranged in the outdoor unit 1, and the refrigerant flow can be switched in the opposite direction by the first four-way valve 35. The first indoor unit 2 has a structure in which a plurality of first heat exchangers 10 mainly composed of sensible heat treatment are provided, and the second indoor unit is provided with a second indoor heat exchanger 13 mainly composed of latent heat treatments. In each heat exchanger, decompression means 9 and 12 are arranged in series, and the second indoor unit 3 is provided with a low-stage compressor 14 and a second four-way valve 36 capable of changing the discharge direction. As a result, the outdoor heat exchanger can be used for both the condenser and the evaporator. Therefore, the heat exchanger disposed in the room can be used for cooling as the evaporator and for heating as the condenser. In the case of a cycle using an ejector as shown in the configuration diagram of FIG. 15, the gas pipe 8 is extended to the second indoor heat exchanger 13 and connected to the anti-expansion valve side, and the first on-off valve is connected to the extension pipe. And a second on-off valve is provided in the suction flow pipe from the second indoor heat exchanger to the ejector 30, and the on-off valve is opened and closed in accordance with the switching of the four-way valve 35 on the heat source side. Heating is also possible. That is, when it is desired to air-condition the room, the already described operation is performed by closing the first on-off valve and opening the second on-off valve. On the other hand, when the heat source side is reversed, when the first on-off valve is opened and the second open / close valve is closed, the refrigerant is supplied to the heat exchangers 13 and 17 in parallel, and both heat exchangers serve as condensers. Operate. The ejector 30 is a pipe through which a refrigerant not related to any operation passes.

  As described above, the air conditioner according to the present invention includes a first compressor whose capacity can be adjusted, an outdoor heat exchanger, an outdoor unit including an outdoor fan, a first decompression unit, and a first indoor heat. Exchanger, first indoor unit including second indoor blower, second decompression means, second indoor heat exchanger, second indoor blower with adjustable air volume, second compression with adjustable capacity In an air conditioner composed of a second indoor unit including an air conditioner and a refrigerant pipe connecting them, an indoor temperature / humidity information detecting means and a refrigerant temperature detecting means flowing through the second indoor heat exchanger And a target refrigerant temperature setting means for controlling the air flow rate of the second indoor fan so that the refrigerant temperature becomes the target refrigerant temperature. In this case, the high-stage compressor as the first compressor Is controlled so as to approach the set temperature which is the target value of the second The low-stage compressor, which is a compressor, is controlled so as to approach the set humidity, which is the target value of indoor humidity, and the blower used for sensible heat main operation, which is the first indoor fan, is linked to the high-stage compressor. The rotation is set based on the operation mode operated by a remote controller or the like, and the operation is performed. However, when there is no such setting, it is preferable to always set the maximum speed. Comfortable indoor air conditioning can be performed with the least amount of energy by such a combination of operations.

  Such a refrigeration cycle apparatus is used for an air conditioner, but when humidity detection is not performed and a humidity set value is not provided, temperature control by a high-stage compressor, dew point temperature control based on evaporation temperature detection, etc. Can be done.

  In addition, since the first indoor heat exchanger surface temperature is operated to be higher than the dew point temperature of the room air, the drain from the heat exchanger having a large surface area is almost eliminated and the drain processing structure is simplified. In addition, the heat exchanger configuration simplifies countermeasures such as inclination and groove structure for dew droplets to flow down the fin surface, and measures against scattering due to wind speed, so that the heat exchanger arrangement and structure can be freely selected. . For example, in the case of a carbon dioxide refrigerant, even if corrugated fins arranged in the lateral direction are used, there is no problem of the flow of dew, so that a more efficient heat exchanger structure can be realized such that it can be used with confidence. In addition, the problem with the rear heat exchanger that forms an inverted V shape with the front heat exchanger in the shape of a dog like a conventional air conditioner indoor unit is that of the treatment of dew droplets adhering to the heat exchanger and the heat exchange capacity. It was an increase.

  FIG. 23 is an explanatory view of the indoor unit structure of an air conditioner in which the heat exchanger performance is greatly improved. The conventional heat exchanger 10 which is the intake air flow side of the indoor unit casing 38 fixed to the wall 37 with respect to the problems of the prior art is configured as a sensible heat main body so that the upper part covers the upper part. Thus, the connection arrangement of the heat transfer tubes, that is, the first indoor unit shown in FIG. 1 is adapted so that the refrigerant having a temperature higher than the dew point temperature flows halfway along the back side. As a result, the sensible heat treatment ability can be greatly increased, and even the shape in which the upper tip hangs down is free from the problem of dew and does not become a problem. On the other hand, the rear heat exchanger is much smaller than the front side, that is, has a latent heat treatment less than 1/3 of the front side. For example, the heat transfer tubes and heat transfer fins have the same shape on both the front side and the rear side as if they were integrated. However, the refrigerant flowing in the rear heat transfer tube is connected to the refrigerant cycle in which the temperature lower than the dew point temperature, that is, the heat exchanger 13 corresponding to the second indoor unit in FIG. Thus, only the rear heat exchanger needs to have a shape that facilitates drainage. In particular, if a rear structure, which is a part for attaching the unit 38 to a wall surface, is used, a deep groove-shaped drain pan 40 is possible, and wastewater treatment to the outside becomes easy. The drain pan 39 on the front side can be used as a temporary water storage measure at high humidity indoors by simply attaching a shallow groove-shaped one. Such a substantially circular heat exchanger on the front side and rear side approaches a single cross-flow fan at the bottom and surrounds it with a shape in which a large space is provided above, thereby greatly increasing the processing capacity. Efficiency can be improved. In addition, the structure can be simplified by combining indoor units into one unit and using one blower. In particular, for a refrigeration cycle using carbon dioxide gas, a structure in which an ejector is provided on the back side of the indoor unit main body, for example, the back of the outlet, etc., can further improve the efficiency with a compact structure. In addition, the ejector 30 and the opening / closing valve 46 for switching between the cooling and heating circuits when the ejector is used can be accommodated in the gap at the back of the casing 38 where the refrigerant pipes 7 and 8 have been provided according to the installation state. An air conditioner that is efficient and uses natural refrigerant can be combined in a small and simple configuration.

Further, according to the present invention, the capacity of the first compressor is controlled based on the deviation between the target temperature and the current temperature, and the capacity of the second compressor is controlled based on the deviation between the target humidity and the current humidity. Not only can individual control with high efficiency be possible, but an effective apparatus is possible because comprehensive indoor heat load processing is performed in the entire refrigeration cycle.

Further, according to the present invention, the second indoor unit includes a third decompression unit and a third indoor heat exchanger, and the second indoor heat exchanger after the indoor air passes through the third indoor heat exchanger. Since it is configured to pass through, it is possible to effectively utilize the processing of the high stage side, that is, the sensible heat load at the time of cooling, and further, it is possible to improve the processing capacity and the efficiency for the heating load. .
is there.

  In addition, since the refrigerant sealed in the refrigerant circuit has an enthalpy difference that can be used at different evaporation temperatures such as carbon dioxide, the refrigerant circuit can be operated with higher efficiency than using a hydrofluorocarbon refrigerant. it can.

  In the air conditioner of the present invention, the first compressor, the heat source side heat exchanger, the first pressure reducing means, and the first load side heat exchanger, which are adjustable or not adjustable in capacity, are sequentially connected to circulate the refrigerant. The first refrigerant cycle, the first refrigerant cycle and the first decompression means and the first load side heat exchanger connected in parallel with the second decompression means, the second load side heat exchanger, and the first refrigerant cycle The second refrigerant cycle that circulates the refrigerant to the second compressor that is operated independently of the compressor, and the air that is air-conditioned in the first load-side heat exchanger and the second load-side heat exchange are the same. Load-side air blowing means that blows out to the air-conditioning region, the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning region, and the surface temperature of the second load-side heat exchanger is The heat treatment capacity of the first load-side heat exchanger is set to the second load so that it is lower than the dew point temperature in the air conditioning area. It is made larger than the heat treatment capability of the heat exchanger.

  In the air conditioner of the present invention, the first compressor whose capacity is adjustable, the heat source side heat exchanger, the first pressure reducing means, and the first load side heat exchanger are sequentially connected to circulate the refrigerant. A first depressurization unit of the refrigerant cycle and the first refrigerant cycle, wherein a second depressurization unit and a second load-side heat exchanger are connected in parallel with the first depressurization unit, and the first depressurization unit An ejector for increasing the refrigerant pressure from the second load-side heat exchanger by using suction by the driving flow at the time of pressure reduction of the means, the first load-side heat exchanger, and the second load-side heat exchange. Load-side air blowing means for blowing air that has been air-conditioned to the same air-conditioning area, and the surface temperature of the first load-side heat exchanger becomes higher than the dew point temperature of the air in the air-conditioning area, The heat of the first load-side heat exchanger is such that the surface temperature of the heat exchanger is lower than the dew point temperature of the air conditioning area. The physical capacity is made larger than the heat treatment capability of the second load-side heat exchanger.

  The air conditioner of the present invention includes an outdoor unit having a capacity-adjustable first compressor, an outdoor heat exchanger, an outdoor blower, a first decompression means, a first indoor heat exchanger, a first The capacity can be adjusted independently of the first indoor unit including the indoor blower, the second decompression means, the second indoor heat exchanger, the second indoor blower capable of adjusting the air volume, and the first compressor. A second indoor unit comprising a second compressor, an outdoor unit, a first indoor unit, and a refrigerant pipe connecting the second indoor unit together on the high pressure side and the low pressure side, Are operated so that the surface temperature of the indoor heat exchanger becomes higher than the dew point temperature of the indoor air.

  The present invention includes an outdoor unit including a first compressor with adjustable capacity, an outdoor heat exchanger, and an outdoor fan, a first decompression unit, a first indoor heat exchanger, and a first indoor fan. A first indoor unit, an expansion power recovery means, a second decompression means, a second indoor heat exchanger, a second indoor blower, a second indoor unit comprising a second compressor, an outdoor unit, The first indoor unit and the second indoor unit are connected by a high pressure side liquid pipe and a low pressure side gas pipe, and the second compressor is driven by the power recovered by the expansion power recovery means by the high pressure side refrigerant. is there.

  In the present invention, one or more first decompression means and first load side heat exchanger, and two or more second decompression means and second load side heat exchangers are provided, respectively, and One load-side heat exchanger and at least one second load-side heat exchanger are provided.

  In the present invention, the first load-side heat exchanger or the first indoor heat exchanger is a sensible heat exchanger whose surface temperature is 10 degrees or more higher than the temperature of air in the air-conditioned region, and the second load-side heat exchange Or the second indoor heat exchanger is a latent heat exchanger whose surface temperature is within 10 degrees from the temperature of air in the air-conditioned area.

  According to the present invention, a third pressure reducing means and a third heat exchanger are provided in parallel with the first pressure reducing means and the first load side heat exchanger or the first indoor side heat exchanger. After passing through the heat exchanger 3, the second load-side heat exchanger or the second indoor heat exchanger is passed.

  The present invention provides an air suction side of the second load side heat exchanger or the second indoor side heat exchanger so as to lower the intake air temperature of the second load side heat exchanger or the second indoor side heat exchanger. Are provided with a sensible heat exchanger different from the first load side heat exchanger or the first indoor side heat exchanger.

  The present invention provides heat between the blowout air of the second load side heat exchanger or the second indoor side heat exchanger and the intake air of the second load side heat exchanger or the second indoor side heat exchanger. A total heat heat exchanger to be exchanged and a sensible heat exchanger such as a heat pipe are provided on the air suction side of the second load side heat exchanger or the second indoor side heat exchanger.

  Further, the present invention switches between the heat source side heat exchanger and the first load side heat exchanger or the first indoor side heat exchanger so that the refrigerant discharge side and the suction side of the first compressor can be switched. A valve is provided. Further, according to the present invention, a second switching valve is provided so as to be switchable between the first compressor side and the heat source side heat exchanger side of the second refrigerant cycle.

  In the present invention, an expander driven by the high pressure of the liquid refrigerant is provided on the high pressure side of the first refrigerant cycle, and a second compressor driven by the expander is provided on the low pressure side of the first refrigerant cycle. It is a thing.

  The present invention comprises refrigerant temperature detecting means for detecting the temperature of the refrigerant flowing through the second load-side heat exchanger or the second indoor heat exchanger, and the second temperature so that the refrigerant temperature becomes the target refrigerant temperature. The air flow rate of the air blowing means for blowing air to the load side heat exchanger or the second indoor heat exchanger is controlled.

  In the present invention, the capacity of the first compressor is controlled based on the deviation between the target temperature and the air temperature in the air-conditioned area, and the capacity of the second compressor is based on the deviation between the target humidity and the air humidity in the air-conditioned area. Be controlled. The intake air of the second load-side heat exchanger or the second indoor-side heat exchanger is intake outside air. Further, the maximum capacity of the first compressor is three times or more the processing heat amount of the second compressor maximum capacity. The heat transfer area of the first load-side heat exchanger or the first indoor heat exchanger is three times or more than the heat transfer area of the second load-side heat exchanger or the second indoor heat exchanger.

  According to the present invention, the first load-side heat exchanger or the first indoor-side heat exchanger, and the second load-side heat exchanger or the second indoor-side heat exchanger are accommodated in one box. Air-conditioning the same air-conditioning basin.

  Further, according to the present invention, the decompression means, the second compressor, the ejector, the on-off valve, and the like are accommodated in a box body such as a unit casing.

  In the air conditioner of the present invention, when the same cooling load is processed with two different evaporating pressures of the low stage and the high stage, the air volume of the low stage side evaporating temperature is adjusted so that the power consumption is minimized. Thus, the entire refrigeration cycle can be operated with high efficiency.

  In addition, since the present invention cools and dehumidifies in the low-stage heat exchanger after pre-cooling the air in the high-stage heat exchanger, the compressor capacity ratio on the low-stage side becomes smaller than that on the high-stage side, and the total Power consumption can be reduced. In addition, the power consumption can be further reduced by introducing the processing air on the lower stage side to be the outside air.

  Further, since the present invention uses carbon dioxide as the refrigerant, it can be operated with higher efficiency than using a hydrofluorocarbon-based refrigerant.

It is a refrigerant circuit figure of the air conditioner which shows Embodiment 1 of this invention. It is a Ph diagram which shows the refrigerating cycle operation | movement of Embodiment 1 of this invention. It is a characteristic view which shows the sensible heat ratio characteristic with respect to the evaporation temperature change of the low stage heat exchanger in Embodiment 1 of this invention. It is a characteristic view which shows the low stage compressor input change with respect to the low stage evaporation temperature of Embodiment 1 of this invention. It is a flowchart which shows the control operation of the air conditioner in Embodiment 1 of this invention. It is a characteristic view which shows the relationship between the low stage evaporation temperature and COP in Embodiment 1 of this invention. It is a refrigerant circuit figure in another embodiment of this invention. It is a Ph diagram which shows the refrigerating cycle operation | movement at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. It is a Ph diagram which shows the refrigerating cycle operation | movement at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. It is another refrigerant circuit figure at the time of enclosing the carbon dioxide refrigerant in Embodiment 1 of this invention. It is explanatory drawing of the air diagram explaining the external air processing state which shows Embodiment 1 of this invention. It is explanatory drawing explaining the pre-cooling and reheating structure of Embodiment 1 of this invention. It is operation | movement explanatory drawing of the pre-cooling and reheating structure in Embodiment 1 of this invention. It is principle explanatory drawing of the ejector of Embodiment 1 of this invention. It is a refrigerating cycle block diagram using the ejector in Embodiment 1 of this invention. It is a Ph diagram in the refrigerating cycle using the ejector of Embodiment 1 of this invention. It is refrigerating cycle operation | movement explanatory drawing using the ejector in Embodiment 1 of this invention. It is another refrigeration cycle block diagram using the ejector in Embodiment 1 of this invention. It is a Ph diagram which shows operation | movement of the refrigerating cycle structure of the air conditioner in Embodiment 1 of this invention. It is a partial refrigeration cycle structure explanatory view using the ejector in Embodiment 1 of this invention. It is structure explanatory drawing of the air conditioner in Embodiment 1 of this invention. It is another refrigeration cycle block diagram in Embodiment 1 of this invention. It is another block diagram of the air conditioner in Embodiment 1 of this invention.

Explanation of symbols

  1 outdoor unit, 2 first indoor unit, 2nd indoor unit, 4 high stage compressor, 5 outdoor heat exchanger, 6 outdoor blower, 7 refrigerant pipe (liquid pipe), 8 refrigerant pipe (gas pipe), DESCRIPTION OF SYMBOLS 9 1st pressure reduction means, 10 1st indoor heat exchanger, 11 1st indoor fan, 12 2nd pressure reduction means, 13 2nd indoor heat exchanger, 14 Low stage compressor, 15 2nd indoor fan, 16 3rd pressure reduction Means, 17 third indoor heat exchanger, 21 expansion power recovery means, 22 expansion section, 23 compression section, 30 ejector, 31 nozzle, 32 mixing section, 33 diffuser section, 34 indoor unit, 35 first four-way valve, 36 Second four-way valve, 37 wall, 38 casing, 39 sensible heat side drain pan, 40 latent heat side drain pan, 41 cross flow fan, 42 outlet, 43 wind guide plate, 44 sensible heat side heat transfer tube, 45 latent heat side heat transfer tube 46 ejector 50 ceiling, 51 ceiling cassette, 52 a blower, 53 a filter, 54 a drain pan, storage box, such as 55-off valve, 56 total heat exchanger.

Claims (17)

An outdoor unit including an outdoor fan that is connected to and condenses the refrigerant compressed by the first compressor capable of capacity adjustment in the heat source side heat exchanger and blows air to the heat source side heat exchanger; A first room having a first indoor blower that is connected and provided to evaporate the refrigerant depressurized by the pressure reducing means in the first indoor heat exchanger and blows air to the first indoor heat exchanger. The refrigerant decompressed by the unit and the second decompression means is evaporated by the second indoor heat exchanger and compressed by the second compressor whose capacity can be adjusted independently of the first compressor. A second indoor unit having a second indoor fan that is connected and has an adjustable air volume to be blown to the second indoor heat exchanger, and a refrigerant condensed in the heat source side heat exchanger of the outdoor unit Is the first indoor unit A high-pressure side liquid pipe connected so as to flow into the second decompression means of the second indoor unit, a refrigerant evaporating in the first indoor heat exchanger of the first indoor unit, and A low-pressure side gas pipe connected so that the refrigerant compressed by the second compressor of the second indoor unit joins and the first compressor of the outdoor unit sucks. The first compressor is operated so that the surface temperature of the indoor heat exchanger becomes higher than the dew point temperature of the indoor air, and the surface temperature of the second indoor heat exchanger becomes lower than the dew point temperature of the indoor air. Thus, the second compressor is operated as described above. An outdoor unit including an outdoor fan that is connected to and condenses the refrigerant compressed by the first compressor capable of capacity adjustment in the heat source side heat exchanger and blows air to the heat source side heat exchanger; The pressure-reduced driving refrigerant flowing into the second decompression means and the pressure-reducing suction refrigerant that has been decompressed and evaporated in the second indoor heat exchanger, sucked by the driving refrigerant in the pressure-increasing means, and pressure-intensified. The first indoor heat exchanger and the first indoor heat exchanger are connected to each other and vaporized by the first indoor heat exchanger, and blown into the first indoor heat exchanger and the second indoor heat exchanger. The first indoor unit having two or more second indoor fans, and the refrigerant condensed in the heat source side heat exchanger of the outdoor unit is the pressure-increasing means of the first indoor unit and the second Connect to the decompression means A high-pressure side liquid pipe that connects the refrigerant evaporating from the first indoor heat exchanger of the first indoor unit so that the first compressor of the outdoor unit sucks the refrigerant. And the pressure raising means sets the evaporation pressure of the second indoor heat exchanger to the suction of the first compressor so that the surface temperature of the second indoor heat exchanger becomes lower than the dew point temperature of the indoor air. An air conditioner characterized by a pressure lower than the pressure. A first compressor having a capacity-adjustable first compressor, a heat source side heat exchanger, an outdoor unit including an outdoor fan, a first pressure reducing means, a first indoor heat exchanger, and a first indoor fan. An indoor unit, an expansion power recovery means, a second decompression means, a second indoor heat exchanger, a second indoor blower, a second indoor unit comprising a second compressor, the outdoor unit, the first 1 indoor unit and the second indoor unit are connected by a high pressure side liquid pipe and a low pressure side gas pipe, and the second compressor is driven by the power recovered by the expansion power recovery means by the high pressure side refrigerant. Air conditioner characterized by. The surface temperature of the second indoor heat exchanger is measured, and the second indoor blower and the second pressure reducing means are adjusted so that the surface temperature of the second indoor heat exchanger is lower than the dew point temperature of the indoor air. The air conditioner according to any one of claims 1 to 3, wherein at least one of the air conditioners is controlled. The first compressor is controlled based on a deviation between a target temperature of the room air and an actual air temperature of the room air, and the second compressor or the booster means the target humidity of the room air and the current of the room air. Or the second indoor heat exchange controlled based on the deviation between the surface temperature or refrigerant temperature of the second indoor heat exchanger and the dew point temperature of the room air. The air conditioner according to any one of claims 1 to 3, wherein the air conditioner is controlled based on a difference between a surface temperature or a refrigerant temperature of the chamber and a set target value. The air conditioner according to any one of claims 1 to 5, wherein a heat transfer area of the first indoor heat exchanger is larger than a heat transfer area of the second indoor heat exchanger. The first indoor heat exchanger is a sensible heat exchanger in which the difference between the surface temperature of the first indoor heat exchanger and the air temperature is within about 10 degrees, and the second indoor heat exchanger is The air conditioner according to any one of claims 1 to 6, wherein the surface temperature of the second indoor heat exchanger is a latent heat exchanger lower by about 10 degrees or more than the air temperature. A third decompression means and a third heat exchanger are provided in parallel with the first decompression means and the first indoor-side heat exchanger, and the indoor air passes through the third heat exchanger and the first The air conditioner according to claim 1, wherein the air conditioner is configured to pass through two indoor heat exchangers. The first indoor heat exchanger is provided on the air intake side of the second indoor heat exchanger so as to lower the intake air temperature of the second indoor heat exchanger. The air conditioner according to any one of 1 to 8. A total heat heat exchanger that exchanges heat between the air blown into the room from the second indoor heat exchanger and the air sucked into the second indoor heat exchanger from the room, or a heat pipe The air conditioner according to any one of claims 1 to 9, wherein another independent sensible heat exchanger is provided on the air suction side of the second indoor heat exchanger. The heat source side heat exchanger is condensed by the heat source side heat exchanger between the first pressure reducer and the second pressure reducer, and then passed through the first pressure reducer and the second pressure reducer. An expander driven by expansion power of refrigerant flowing into the first indoor heat exchanger and the second indoor heat exchanger is connected, and the expander is connected to a low pressure side of the second indoor heat exchanger. The air conditioner according to claim 1, 3, or 6, wherein the second compressor is connected to compress the refrigerant evaporating from the second indoor heat exchanger. Refrigerant temperature detecting means for detecting the temperature of the refrigerant flowing through the second indoor heat exchanger, and blowing air to the second indoor heat exchanger so that the refrigerant temperature is lower than the dew point temperature of the indoor air. The air conditioner according to any one of claims 1 to 11, wherein the air volume of the second indoor fan is controlled. The air conditioner according to any one of claims 1 to 9, 11 or 12, wherein the intake air of the second indoor heat exchanger is outside air. The air conditioner according to any one of claims 1 and 3 to 13, wherein the maximum capacity of the first compressor is three times or more the amount of heat of treatment of the maximum capacity of the second compressor. . The air conditioning according to any one of claims 1 to 14, wherein a heat transfer area of the first indoor heat exchanger is three times or more a heat transfer area of the second indoor heat exchanger. Machine. 2. The opening degree of the first decompression unit and the second decompression unit is adjusted so that a refrigerant superheat degree at a heat exchanger outlet connected to one end of the first decompression unit and the second decompression unit becomes a predetermined value. The air conditioner according to any one of 1 to 15. If the evaporation temperature of the refrigerant sealed in the refrigerant circuit formed by connecting the outdoor unit, the first indoor unit, and the second indoor unit by the high-pressure side liquid pipe and the low-pressure side gas pipe is different, enthalpy The air conditioner according to any one of claims 1 to 16, wherein the air conditioner is a refrigerant having a difference in difference.

Images