JP7142785B2 - Air conditioner - Google Patents

Description

The present invention relates to an air conditioner.

While using various alternative refrigerants to replace HCFC (Hydro Chloro Fluoro Carbon) refrigerant, which has the potential to deplete the ozone layer, it has a coefficient of performance in actual equipment that is comparable to HCFC refrigerants, and can safely use alternative refrigerants as operating refrigerants. There is a demand for an air conditioner that can

Japanese Patent Laying-Open No. 7-269964 (Patent Document 1) discloses an example of such an air conditioner. This air conditioner has a flow path A containing a refrigerant a and a flow path B containing a refrigerant b, and an intermediate heat exchanger in which the refrigerants a and b exchange heat with each other.

JP-A-7-269964

In recent years, the prevention of global warming has become an urgent issue, and the Kigali Amendment to the Montreal Protocol and the F-gas regulations in Europe have been enacted. It has been demanded. HFC (Hydro Fluoro Carbon) refrigerants such as R410A have mainly been used in vapor compression refrigeration cycles, which are heat source generators for air conditioners. For this reason, HC (Hydrocarbon) refrigerants such as propane, HFO (Hydro Fluoro Olefin) refrigerants, mixed refrigerants containing these refrigerants as main components, and the like have been proposed as alternative refrigerants to HFC refrigerants.

Propane is a highly flammable refrigerant, so when used indoors, international standards (IEC, etc.) and domestic standards (ISO, High Pressure Gas Safety Law, etc.) impose safety measures and limits on the amount of refrigerant charged. For this reason, the primary circuit that functions as a heat source is placed outdoors, and a secondary circuit that uses a heat transfer medium that is separate from the primary circuit is configured indoors. Indirect air conditioning systems have been developed that deliver to exchangers.

Generally, in indirect air conditioning systems, water is used as the heat-carrying medium on the secondary circuit side and a pump is used as the drive. When the sensible heat of water is used, the liquid flow rate is larger than when the latent heat of the refrigerant is used, and the size of the pump and liquid piping for conveying water is increased, and the power is also increased. For this reason, it was difficult to replace the refrigerant pipes of the direct air conditioner using the existing refrigerant with the water pipes. In the first place, water is unsuitable as a heat transfer medium for facilities where it is prohibited to bring water pipes indoors, such as data centers where many large computers are installed.

In addition, in the case of an air conditioner using R32 as a refrigerant in the primary circuit, the theoretical coefficient of performance (hereinafter referred to as theoretical COP) determined from the thermophysical values is high, but the GWP value is 675, which is higher than the international regulatory standards. It gets pretty big. In addition, since R32 is weakly flammable, safety measures were required when used indoors. Therefore, in order to improve the GWP and combustibility characteristics while keeping the theoretical COP at the same level as R32, a refrigerant mixed with an HFO-based refrigerant having low GWP characteristics has been developed.

It is conceivable to employ the configuration disclosed in Japanese Patent Application Laid-Open No. 7-269964 for an indirect air-conditioning system using such a refrigerant. Japanese Patent Application Laid-Open No. 7-269964 discloses an indirect air-conditioning system in which the primary circuit and the secondary circuit are filled with different Freon refrigerants, and a four-way valve in the primary circuit switches between cooling operation and heating operation. there is Important points for ensuring high efficiency and reliability of this air conditioner include ensuring a sufficiently large degree of supercooling in the primary circuit and preventing gas refrigerant from being sucked into the secondary circuit pump. However, the configuration disclosed in Japanese Patent Application Laid-Open No. 7-269964 does not have a circuit configuration that considers these important points.

SUMMARY OF THE INVENTION The present invention has been made to solve the above problems, and an air conditioning system comprising a heat source side first refrigerant circuit containing a first refrigerant and a room side second refrigerant circuit containing a second refrigerant. An object of the present invention is to provide an air conditioner capable of ensuring high efficiency and reliability.

The present disclosure relates to an air conditioner comprising an intermediate heat exchanger configured to exchange heat between a first refrigerant and a second refrigerant. The air conditioner includes a first refrigerant circuit containing a first refrigerant and a second refrigerant circuit containing a second refrigerant. The first refrigerant circuit includes a compressor that compresses a first refrigerant in a gaseous state, an outdoor heat exchanger that exchanges heat between the first refrigerant and the outside air, an expansion device that decompresses and expands the first refrigerant, and a first flow path through which the first refrigerant passes in the intermediate heat exchanger. The second refrigerant circuit includes a pump that pressurizes and conveys the second refrigerant in a liquid state, a second flow path through which the second refrigerant passes, and heat exchange between the second refrigerant and room air in the intermediate heat exchanger. and a first tank for degassing provided in the middle of the first pipe in the second refrigerant circuit. The first pipe is a pipe connected to the pump and through which the second refrigerant sucked by the pump passes.

According to the air conditioner of the present disclosure, since the tank for degassing is provided in the suction portion of the pump of the second refrigerant circuit, suction of gas refrigerant into the pump is suppressed. Therefore, high efficiency and reliability can be ensured in the air conditioner.

4 is a refrigerant circuit diagram during cooling operation of the air conditioner according to Embodiment 1. FIG. FIG. 2 is a ph diagram of the vapor compression refrigeration cycle of FIG. 1; 2 is an enlarged view showing the vicinity of the pump 23 in FIG. 1 during cooling operation; FIG. 4 is a refrigerant circuit diagram during heating operation of the air conditioner according to Embodiment 1. FIG. FIG. 5 is a ph diagram of the vapor compression refrigeration cycle of FIG. 4; 5 is an enlarged view showing the vicinity of the pump 23 in FIG. 4 during heating operation; FIG. 2 is a diagram showing an example of a configuration of a pump arranged in a second refrigerant circuit 3 of FIG. 1; FIG. FIG. 3 is a plan view of one impeller of the Wesco pump. FIG. 9 is a sectional view along section IX-IX in FIG. 8; It is the perspective view which expanded and showed the blade|wing part of the impeller. 4 is a diagram showing how to combine the forward direction and the reverse direction of the direction of the first refrigerant by the four-way valve 12 of the first refrigerant circuit 2 and the forward direction and the reverse direction of the rotation direction of the pump 23 of the second refrigerant circuit 3. FIG. 4 is a flowchart for explaining refrigerant circulation direction switching control executed by the control device 100. FIG. FIG. 4 is a diagram for explaining the GWP total amount value of Embodiment 1 and the GWP total amount value of a comparative example; FIG. 7 is a refrigerant circuit diagram during cooling operation of the air conditioner according to Embodiment 2; FIG. 7 is a refrigerant circuit diagram during heating operation of the air conditioner according to Embodiment 2;

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. A plurality of embodiments will be described below, but appropriate combinations of the configurations described in the respective embodiments have been planned since the filing of the application. The same or corresponding parts in the drawings are denoted by the same reference numerals, and the description thereof will not be repeated.

Embodiment 1.
FIG. 1 is a refrigerant circuit diagram during cooling operation of an air conditioner according to Embodiment 1. FIG. FIG. 2 is a ph diagram of the vapor compression refrigeration cycle of FIG.

As shown in FIG. 1 , the air conditioner 1 includes a first refrigerant circuit 2 , a second refrigerant circuit 3 and a control device 100 . The first refrigerant circuit 2 corresponds to an "outdoor cycle", a "heat source side cycle", or a "primary circuit". The second refrigerant circuit 3 corresponds to an "indoor cycle", a "utilization side cycle", or a "secondary circuit".

The first refrigerant circuit 2 includes a compressor 10, a four-way valve 12, an outdoor heat exchanger 13, an expansion device 24, a first flow path H1 of an intermediate heat exchanger 22, and a pipe 7 as main elements.

The second refrigerant circuit 3 includes, as main elements, the second flow path H2 of the intermediate heat exchanger 22, the pump 23, the first tank 17, the second tank 18, the indoor temperature control units 30 and 40, and the pipes 11L and 11G. include. The indoor temperature control units 30 and 40 are connected in parallel between the pipe 11L and the pipe 11G.

The indoor temperature control unit 30 includes an indoor heat exchanger 31, a fan (not shown) for sending indoor air to the indoor heat exchanger 31, and a flow control valve 33 for adjusting the flow rate of the second refrigerant. The indoor heat exchanger 31 exchanges heat between the second refrigerant and the indoor air.

The indoor temperature control unit 40 includes an indoor heat exchanger 41, a fan (not shown) for sending indoor air to the indoor heat exchanger 41, and a flow control valve 43 for adjusting the flow rate of the second refrigerant. The indoor heat exchanger 41 exchanges heat between the second refrigerant and the indoor air.

In this embodiment, an air conditioner having two indoor temperature control units is taken as an example, but the number of indoor temperature control units may be any number.

The first refrigerant circuit 2 seals the first refrigerant to form a closed circuit. The first refrigerant used in Embodiment 1 is, for example, an HFC refrigerant such as R32 or an HC refrigerant such as propane (R290). The second refrigerant circuit 3 seals the second refrigerant to form a closed circuit. The second refrigerant used in Embodiment 1 is, for example, a mixed refrigerant such as R466A containing CF ₃ I.

One of the features of this embodiment is that at least one of the first refrigerant and the second refrigerant has a global warming potential smaller than that of R32, and the second refrigerant has a lower combustion limit concentration higher than that of the first refrigerant. , a combination of a first refrigerant and a second refrigerant.

First, the 1st refrigerant circuit 2 is demonstrated. The compressor 10 sucks and compresses a low-temperature, low-pressure gaseous first refrigerant and discharges a high-temperature, high-pressure gaseous first refrigerant. The four-way valve 12 is provided downstream of the discharge port of the compressor 10 and configured to switch the flow path of the refrigerant discharged from the compressor 10 between cooling operation and heating operation. During cooling operation in which the refrigerant is circulated in the direction indicated by the dashed arrow in FIG. On the other hand, during heating operation in which the refrigerant is circulated in the direction indicated by the solid arrow in FIG.

The outdoor heat exchanger 13 functions as a condenser that condenses and liquefies the high-temperature, high-pressure gaseous first refrigerant during cooling operation, and releases heat to the outside air. Conversely, during heating operation, the outdoor heat exchanger 13 functions as an evaporator that evaporates the low-temperature, low-pressure liquid first refrigerant to take heat from the outside air. The expansion device 24 decompresses and expands the refrigerant passing through it to a low-temperature, low-pressure refrigerant. An electronic expansion valve, for example, can be used as the expansion device 24 .

The intermediate heat exchanger 22 is configured to exchange heat between the first refrigerant circulating in the first refrigerant circuit 2 and the second refrigerant circulating in the second refrigerant circuit 3 . The intermediate heat exchanger 22 is formed with a first coolant gap channel and a second coolant gap channel. A low-temperature side entrance and a high-temperature side entrance are provided in the first coolant clearance channel, and a low-temperature side entrance and a high-temperature side entrance are provided in the second coolant clearance channel. If the low temperature side entrance and the high temperature side entrance/exit of the second refrigerant clearance channel are appropriately arranged with respect to the high temperature side entrance and the low temperature side entrance of the first refrigerant clearance channel, the first refrigerant and the second refrigerant flow in opposite directions. Become.

It is preferable to use a plate heat exchanger as the intermediate heat exchanger 22 because the plate heat exchanger is generally superior in terms of cost and efficiency. For example, in a plate heat exchanger, a plurality of rectangular corrugated plates are laminated to form a plurality of rectangular gaps. The plurality of rectangular gaps alternately form the first coolant gap channel and the second coolant gap channel in the stacking direction.

In the first refrigerant circuit 2 , a four-way valve 12 is provided in a pipe that discharges the first gaseous refrigerant from the compressor 10 . The four-way valve 12 guides the first refrigerant from the discharge port of the compressor 10 to the outdoor heat exchanger 13 side during cooling operation, and guides the first refrigerant to the intermediate heat exchanger 22 side during heating operation. switch the circulation direction of

On the other hand, in the second refrigerant circuit 3, the pump 23 is configured to switch the rotation direction between the first rotation direction and the opposite second rotation direction. The pump 23 guides the liquid second refrigerant from the pump 23 to the indoor heat exchangers 31 and 41 during cooling operation, and directs the liquid second refrigerant from the pump 23 to the second flow path of the intermediate heat exchanger 22 during heating operation. The circulation direction of the second refrigerant is switched so as to lead to H2.

In this manner, the second refrigerant circulating in the second refrigerant circuit 3 is cooled by the first refrigerant circulating in the first refrigerant circuit 2 during cooling operation. On the other hand, during heating operation, the first refrigerant circulating in the first refrigerant circuit 2 heats the second refrigerant circulating in the second refrigerant circuit 3 .

The pipe 7 connects the first flow path H1 of the compressor 10, the four-way valve 12, the outdoor heat exchanger 13, the expansion device 24, and the intermediate heat exchanger 22, and allows the first refrigerant to circulate between these components. It is intended to Also, the intermediate heat exchanger 22, the pump 23, the first tank 17, and the second tank 18 form the intermediate heat exchange unit 20 and are normally arranged on the outdoor side.

The pump 23 is suitable as means for boosting and conveying liquid refrigerant, and is of a type capable of forward and reverse rotation. Such types of pumps include, for example, a Wesco type (also called a vortex type or a regenerative type), a screw type, and the like. In the air conditioner 1 of Embodiment 1, the pump 23 on the second refrigerant circuit 3 side raises the pressure of the second refrigerant in the liquid state at the same volume, so the theoretical enthalpy change amount can be small. Therefore, the power required in the pump 23 can be kept small.

The indoor heat exchangers 31 and 41 exchange heat between the second refrigerant passing through them and indoor air. The indoor heat exchangers 31 and 41 function as so-called evaporators during cooling operation, and function as gas coolers during heating operation.

The pipes 11L, 11G connect the intermediate heat exchanger 22, the pump 23, and the indoor heat exchangers 31, 41. A second refrigerant circuit 3 through which the second refrigerant can circulate is formed by connecting these components with pipes 11L and 11G. In addition, flow control valves 33 and 43 are provided in the pipes 11L connected to the inlet sides of the indoor heat exchangers 31 and 41, respectively. By adjusting the opening degrees of the flow control valves 33 and 43, the flow rate of the second refrigerant passing through the indoor heat exchangers 31 and 41 installed for each room from the pump 23 can be adjusted.

Under the cooling conditions in FIG. 1, the pipe 11L from the point A2 near the outlet of the intermediate heat exchanger to the inlets of the indoor heat exchangers 31 and 41 via the pump 23 is a pipe mainly for circulating liquid refrigerant. A pipe 11G extending from the outlets of the indoor heat exchangers 31 and 41 to the inlet E2 of the intermediate heat exchanger is a pipe mainly for circulating gas refrigerant.

For each room, an indoor temperature control unit 30 including an indoor heat exchanger 31 and a flow control valve 33 and an indoor temperature control unit 40 including an indoor heat exchanger 41 and a flow control valve 43 are arranged. Indoor air temperature can be adjusted.

Switching of the refrigerant circulation direction performed in first refrigerant circuit 2 and second refrigerant circuit 3 is performed by control device 100 .

Controller 100 includes processor 102 and memory 103 . The memory 103 includes, for example, a ROM (Read Only Memory), a RAM (Random Access Memory), and a flash memory. The flash memory stores an operating system, application programs, and various data. Processor 102 controls the overall operation of air conditioner 1 . Control device 100 shown in FIG. 1 is implemented by processor 102 executing an operating system and application programs stored in memory 103 . Various data stored in the memory 103 are referred to when executing the application program.

If the control device 100 receives a signal from a remote controller, it may further include a receiving device. When there are a plurality of indoor temperature control units, such a receiving device is provided corresponding to each of the plurality of indoor units.

Note that when the control device is divided into a plurality of control units, each of the plurality of control units includes a processor. In such a case, a plurality of processors work together to control the air conditioner 1 as a whole.

The control device 100 controls the compressor 10, the expansion device 24, the pump 23, the flow control valves 33 and 43, and the fan (not shown) according to the outputs of the pressure sensor, temperature sensor, and the like.

The control device 100 switches the circulation direction of the first refrigerant in the first refrigerant circuit 2 by the four-way valve 12 between the refrigerant operation and the heating operation. In conjunction with this, the control device 100 controls the second refrigerant circuit so that the second refrigerant exchanges heat with the first refrigerant in a counterflow manner in the intermediate heat exchanger 22 and becomes supercooled at the suction port of the pump 23. The rotation direction of the pump 23 of No. 3 is switched.

(cooling operation control)
Now, when the air conditioner 1 configured as shown in FIGS. 1 and 2 is set to the cooling operation mode, the compressor 10 keeps the outlet temperature at a point A1 near the outlet of the intermediate heat exchanger 22 constant. Based on the signal from the temperature sensor 63 located near the outlet of the intermediate heat exchanger 22, the controller 100 controls the rotation speed of the compressor 10 so as to maintain the temperature.

In addition, based on the signal from the temperature sensor 63, the control device 100 controls the rotation speed of the pump 23 so as to keep the degree of subcooling constant near the outlet of the intermediate heat exchanger 22 of the second refrigerant indicated by the point A2. do.

Also, from the temperature sensors 68, 69 arranged near the exits of the indoor heat exchangers 31, 41 so as to keep constant the degree of superheat at the positions indicated by points D21, D22 near the exits of the indoor heat exchangers 31, 41 The control device 100 controls the opening degrees of the flow control valves 33 and 43 based on the signals of .

That is, when the degrees of superheat at points D21 and D22 detected by the temperature sensors 68 and 69 are higher than the "target degree of superheat" preset as a target, the opening degrees of the flow control valves 33 and 43 are increased. , when the degree of superheat is low, the opening degrees of the flow control valves 33 and 43 are reduced. In this manner, the degree of opening of the flow control valves 33, 43 is controlled so that the degree of superheat near the outlets of the indoor heat exchangers 31, 41 is kept constant.

In the intermediate heat exchanger 22, the first refrigerant in a low-pressure liquid state flows in from a point D1 near the inlet, undergoes a phase change while taking heat from the second refrigerant, and flows out in a low-pressure gas state to A1 near the outlet. On the other hand, the second refrigerant flows in from a point E2 near the inlet in a low-pressure gas state, undergoes a phase change while absorbing heat from the first refrigerant, flows out in a low-pressure liquid state to a point A2 near the outlet, and is sucked into the pump 23. be. At this time, since the first refrigerant and the second refrigerant exchange heat in a countercurrent manner, highly efficient heat exchange can be realized. Here, in order to keep the efficiency and reliability of the pump 23 at a high level, it is important to keep the second refrigerant sucked into the pump 23 in a liquid state.

However, the first refrigerant at point D1 is not in a supercooled state but in a gas-liquid mixed state, which is unsuitable for increasing the degree of supercooling of the second refrigerant that exchanges heat. In this state, for example, point A2 in FIG. 2 is drawn toward point D1 and enters the two-phase region, and the second refrigerant is likely to be sucked into pump 23 in a gas-mixed state.

In Embodiment 1, in order to avoid this, the first tank 17 for degassing is arranged in the piping on the suction side of the pump 23 of the second refrigerant circuit 3 .

FIG. 3 is an enlarged view showing the vicinity of the pump 23 in FIG. 1 during cooling operation. Referring to FIG. 3 , a first tank 17 for degassing is arranged between pipe 11L1 and pipe 11L2 which are on the absorption side of pump 23 during cooling operation of second refrigerant circuit 3 . Further, a second tank 18 for degassing is arranged between the pipe 11L3 and the pipe 11L4 on the discharge side of the pump 23 during the cooling operation of the second refrigerant circuit 3 .

The first tank 17 includes a space 17A for storing liquid refrigerant on the bottom side of the tank housing 17C and a space 17B for storing gas refrigerant on the top side of the tank housing 17C during cooling operation.

The second tank 18 includes a space 18A for storing liquid refrigerant on the bottom side of the tank housing 18C and a space 18B for storing gas refrigerant on the top side of the tank housing 18C during heating operation.

The pump 23 includes a motor 23A, a pump housing 23B, and a vortex impeller 23D. A space that accommodates the impeller 23D inside the pump section housing 23B is a cylinder 23C. The pump 23 drives the rotary shaft of the impeller 23D.

During cooling operation, when the secondary refrigerant mixed with gas that has not reached a saturated state flows from the intermediate heat exchanger 22 into the first tank 17 for gas venting through the pipe 11L, the space on the lower bottom side of the container Liquid refrigerant is stored in 17A, and gas refrigerant is accumulated in space 17B on the upper surface side of the container. As a result, only the liquid refrigerant is sucked into the cylinder 23C of the pump 23 through the pipe 11L2. The liquid refrigerant pressurized by the impeller 23D in the cylinder 23C is discharged through the pipe 11L3. Here, the second tank 18 for venting gas is also provided on the discharge side during cooling. passes through the second tank 18 for degassing and the pipe 11L4 as it is as a liquid refrigerant.

As described above, each device can be efficiently operated, and the suction of gas refrigerant in the pump 23 is also suppressed, so that the coefficient of performance COP corresponding to the energy consumption efficiency of the air conditioner 1 as a whole can be maintained at a high level. can be done.

(Heating operation control)
4 is a refrigerant circuit diagram during heating operation of the air conditioner according to Embodiment 1. FIG. FIG. 5 is a ph diagram of the vapor compression refrigeration cycle of FIG.

When the air conditioner 1 is set to the heating operation mode, the temperature at the point D1 near the outlet of the intermediate heat exchanger 22 is kept constant so that the outlet temperature at the position indicated by the point D1 near the outlet of the intermediate heat exchanger 22 is kept constant. The controller 100 controls the rotation speed of the compressor 10 based on the signal from the temperature sensor 61 arranged.

Further, in the second refrigerant circuit 3, the opening degrees of the flow control valves 33 and 43 are adjusted so that the outlet temperatures at the points C21 and C22 near the outlets of the two indoor heat exchangers 31 and 41 are maintained at the respective target temperatures. controlled. The controller 100 controls the opening degrees of the flow control valves 33 and 43 based on signals from the temperature sensors 75 and 76 arranged near the outlets of the indoor heat exchangers 31 and 41 .

By controlling as described above, the air conditioner 1 can be operated at an efficient and optimal operating point according to the load on each of the indoor heat exchangers 31 and 41 .

At this time, it is important to ensure that the degree of subcooling at point D1 of the first refrigerant circuit is ensured, and that the second refrigerant sucked into the pump 23 is brought into a liquid state. To keep.

In the intermediate heat exchanger 22, the first refrigerant flows in a high-pressure gas state from a point A1 near the inlet, undergoes a phase change while being deprived of heat by the second refrigerant, and flows out in a high-pressure liquid state to a point D1 near the outlet. On the other hand, the second refrigerant in a high-pressure liquid state flows in from a point A2 near the inlet, undergoes a phase change while taking heat from the first refrigerant, and flows out in a high-pressure gas state to a point E2 near the outlet. At this time, the first refrigerant and the second refrigerant exchange heat in a countercurrent manner, so that the degree of subcooling can be reliably ensured, and the refrigeration cycle efficiency of the primary circuit can be maintained.

FIG. 6 is an enlarged view showing the vicinity of the pump 23 in FIG. 4 during heating operation. Referring to FIG. 6, in the first embodiment, the rotation direction is switched so that the pump rotates in the reverse direction during the heating operation. Therefore, in the second refrigerant circuit 3, the direction of refrigerant circulation during heating operation is opposite to that during cooling operation.

When the secondary refrigerant mixed with gas that has not reached a saturated state flows from the indoor heat exchangers 31 and 41 during heating operation into the second tank 18 for gas venting through the pipe 11L4, the space on the bottom side of the container Liquid refrigerant is stored in 18A, and gas refrigerant is accumulated in space 18B on the upper surface side of the container. Therefore, only the liquid refrigerant is sucked into the cylinder 23C of the pump 23 through the pipe 11L3. The liquid refrigerant pressurized by the impeller 23D in the cylinder 23C is discharged through the pipe 11L2. Since the first tank 17 for degassing is on the discharge side during heating, the gas refrigerant rarely accumulates in the space 17B on the upper surface side of the container. pass through.

As shown in the ph diagram of FIG. 5, the outlet point D1 of the primary refrigerant of the intermediate heat exchanger 22 is usually controlled near the saturated liquid point, so the inlet point A2 of the secondary refrigerant of the intermediate heat exchanger 22 Points are drawn near the saturated liquid point. Therefore, it is difficult for the dryness at the point A2 of the inlet of the second refrigerant to increase. Therefore, since the degassing effect during heating may be auxiliary, the internal volume of the first tank 17 may be smaller than that of the second tank 18 . As described above, in the present embodiment, two large and small degassing tanks, that is, the first tank 17 and the second tank 18 are arranged before and after the pump 23 .

By controlling as described above, each device can operate efficiently even during heating operation, so that the coefficient of performance COP corresponding to the energy consumption efficiency of the air conditioner 1 as a whole can be further improved.

(Configuration of pump 23)
FIG. 7 is a diagram showing an example of the configuration of a pump arranged in the second refrigerant circuit 3 of FIG. 1. As shown in FIG. FIG. 7 shows the configuration of the cylinder 23C of the vortex pump. An impeller 23D is arranged in the cylinder 23C. The central shaft of the impeller 23D is attached to the drive shaft of the motor. The impeller 23D is rotatable in a first rotation direction R1 indicated by a solid line and a second rotation direction indicated by a broken line. When the impeller 23D rotates in the first rotation direction R1 (counterclockwise) indicated by the solid arrow, the liquid refrigerant sucked from the pipe 72 flows counterclockwise through the flow path 23E of the cylinder 23C while increasing the pressure. It flows out from 73. Conversely, when the impeller 23D rotates in the second rotation direction R2 (clockwise) indicated by the dashed arrow, the liquid refrigerant sucked from the pipe 73 flows clockwise through the flow path 23E of the cylinder 23C while increasing the pressure. It flows out from the pipe 72 .

FIG. 8 is a plan view of one impeller of the Wesco pump. 9 is a cross-sectional view taken along the line IX-IX in FIG. 8. FIG. FIG. 10 is a perspective view showing an enlarged blade portion of the impeller.

The blades of the vortex impeller have a shape in which grooves are dug at regular intervals around a rotating disk. The shape of one sheet is a shape with a tapered tip as shown in FIG. Specifically, each blade that rotates around the drive shaft of the vortex impeller has a first corner portion on the front end side and a second corner portion on the rear end side with respect to the first rotation direction R1 indicated by the solid line arrow in FIG. have The radius of curvature of the second corner on the rear end side is larger than the radius of curvature of the first corner on the front end side.

By increasing the radius of curvature of the corners on the rear end side of the blades rotating in the first rotation direction R1, the lift can be further increased.

In the first embodiment, the impeller 23D is rotated in the first rotation direction R1 during the heating operation, and rotated in the second rotation direction R2, which is the opposite direction during the cooling operation, so that the pump lift during the heating operation is larger than that during the cooling operation. increased.

The head necessary for the cooling operation is the difference in pressure between points A2 and B2 in FIG. 2 (ΔPab=P(B2)-P(A2)). On the other hand, during the heating operation, the pressure difference between points B2 and A2 in FIG. 5 (ΔPba=P(A2)−P(B2)) is required.

The differential pressure ΔPab between points A2 and B2 is equal to the sum of the pressure loss in the secondary liquid refrigerant pipe from point B2 to points C21 and C22 and the pressure loss in the secondary gas refrigerant pipe from points D21 and D22 to point E2. This pressure loss is proportional to the square of (refrigerant circulation amount G) and inversely proportional to the refrigerant density.

Since the second refrigerant is used indoors, the optimum living environment temperature is in the range of about 10°C to 25°C. Since the second refrigerant is controlled within the range of about 15° C. to 30° C. during heating and in the range of about 5° C. to 20° C. during cooling, variations in the density of the second refrigerant are small.

On the other hand, the capacity (=refrigerant circulation amount×enthalpy difference) required for the air conditioner changes depending on the season. The maximum capacity is required during heating operation, and the refrigerant circulation amount is also maximum. Therefore, the maximum lift required by the pump is expected to occur at heating conditions where condensing temperature is low and maximum capacity is required. Therefore, in the first embodiment, as described with reference to FIGS. 7 to 10, the pump 23 is designed to rotate in the first rotation direction R1 during the heating operation to exhibit the maximum pumping capacity.

(Switching control of refrigerant circulation direction)
11 shows the forward and reverse directions of the first refrigerant flowing through the four-way valve 12 of the first refrigerant circuit 2 and the forward and reverse directions of the second refrigerant flowing through the pump 23 of the second refrigerant circuit 3. It is a figure which shows how to combine.

In FIG. 11, regarding the four-way valve 12, the direction in which the first refrigerant discharged from the compressor 10 flows toward the outdoor heat exchanger 13 is defined as the "forward direction", and the first refrigerant discharged from the compressor 10 flows in the middle direction. The direction of flow toward the heat exchanger 22 is described as "reverse". As for the direction of the second refrigerant flowing through the pump 23, the direction in which the refrigerant sent from the pump 23 flows toward the indoor heat exchangers 31 and 41 is described as the "forward direction", and the refrigerant sent from the pump 23 flows toward the intermediate heat exchanger 22 is described as the "reverse direction". 9, when the pump 23 is rotated in the "first direction of rotation," the second refrigerant flows in the "reverse direction," and when the pump 23 is rotated in the "second direction of rotation," , the second refrigerant flows in the “forward direction”.

In the intermediate heat exchanger 22, a combination of "forward direction" and "forward direction" is selected in the cooling operation so that the first refrigerant and the second refrigerant flow countercurrently, and in the heating operation, the "reverse direction" is selected. ' and 'reverse' are selected. A combination in which one is in the "forward direction" and the other is in the "reverse direction" results in parallel flow, which is inefficient and therefore prohibited from use.

The method of controlling the operation while ensuring the necessary capacity under the temperature environment conditions for each of the heating operation and the cooling operation is basically the same as general control. However, in the embodiment of the present invention, the operation is executed with particular attention to the following points 1) to 3) for highly efficient operation.

In the first refrigerant circuit in which the first refrigerant circulates, 1) the degree of subcooling (SC: subcool) of the first refrigerant at the evaporator outlet is ensured, and 2) the degree of superheat (SH : Superheat) is ensured.

On the other hand, in the second circuit in which the second refrigerant circulates, 3) the second refrigerant sucked into the pump 23 is operated so as to be maintained in a liquid phase state. This is important not only for efficiency but also for ensuring pump reliability.

FIG. 12 is a flowchart for explaining switching control of the refrigerant circulation direction executed by the control device 100 . 1 and 12, in step S1, control device 100 determines whether the operation mode designated by a remote control or the like is cooling or heating.

When cooling is determined in step S1, the control device 100 sets the four-way valve 12 so that the first refrigerant circulates in the forward direction in step S2. Furthermore, in step S3, the control device 100 rotates the pump 23 so that the second refrigerant circulates in the forward direction. By controlling in this way, the first refrigerant and the second refrigerant circulate as indicated by the dashed arrows in FIG. and the second refrigerant flowing through the second refrigerant flow countercurrently.

On the other hand, when it is determined in step S1 that heating is performed, the control device 100 sets the four-way valve 12 so that the first refrigerant circulates in the reverse direction in step S4. Furthermore, in step S5, the control device 100 rotates the pump 23 so that the second refrigerant circulates in the reverse direction. By controlling in this way, the first refrigerant and the second refrigerant circulate as indicated by the solid line arrows in FIG. and the second refrigerant flowing through the second refrigerant flow countercurrently.

In the example described above, the direction of circulation of the first refrigerant in the first refrigerant circuit 2 is switched by the four-way valve 12 arranged after the discharge of the compressor 10 as means for switching between the cooling operation and the heating operation. On the other hand, in the second refrigerant circuit 3, by switching the rotation direction of the pump 23, the circulation direction of the second refrigerant is switched. By performing this in conjunction with each other, the second refrigerant sucked by the pump 23 is maintained while the first refrigerant and the second refrigerant flowing through the intermediate heat exchanger 22 are heat-exchanged in a countercurrent manner. It can be kept in a liquid state.

(Explanation of refrigerant types that reduce the total GWP value)
In the vapor compressor refrigerating cycle of air conditioners, first, low GWP refrigerants are being promoted while still using HFC refrigerants, which are excellent in energy-saving performance. For example, conversion from R410A with GWP of 2090 to R32 with GWP of 675, for example. However, there is a considerable gap between the reduction target of the total GWP value by international regulations.

Embodiment 1 describes an example in which a highly flammable refrigerant (classification A3) is used as the first refrigerant and a nonflammable refrigerant is used as the second refrigerant. The GWP is as small as 4 when propane (R290) is used as the first refrigerant. However, the use of highly flammable refrigerants indoors is restricted by laws and regulations.

Next, as the second refrigerant, which is a heat transfer medium to the indoor side, nonflammable and nontoxic R466A (mixed refrigerant of R32, R125, and CF ₃ I) was used in consideration of the case of a building that dislikes water. The GWP of R466A is 733, which is the same level as R32 and about 1/3 of R410A.

For example, the refrigerant charge volume of 10 L is divided into 5 L in the first refrigerant circuit 2 and 5 L in the second refrigerant circuit 3 . Of the 5L in the first refrigerant circuit 2, the intermediate heat exchanger 22 has 1L. Further, it is assumed that 1 L of the 5 L of the second refrigerant circuit 3 is enclosed in the intermediate heat exchanger 22, 1 L is enclosed in the total of the indoor temperature control units 30, 40, and 50, and 1 L is enclosed in the piping 11 other than that.

Below, the GWP total amount value (=Σ(refrigerant volume×density)) of each enclosed refrigerant is compared. Here, the effect of refrigerant density is represented by "×Σρ". Refrigerant density varies from a saturated liquid state to a saturated gas state in the evaporative heat exchanger and the condensing heat exchanger. Here, it is assumed that the first refrigerant has a condensation temperature (CT) of 45°C and an evaporation temperature (ET) of 10°C. ) and the average value ρave was used.

The GWP total amount value of each refrigerant was calculated from GWP total amount value=GWP×refrigerant volume×average density ρave. In the present embodiment, the specific GWP total value was calculated based on the R32 refrigerant direct expansion type refrigerant specific GWP total value, and the reduction effect depending on the type of refrigerant in the secondary circuit was shown.

On the other hand, the pressure loss and refrigerant circulation amount during turbulent flow are expressed by the following equations.
Pressure loss = Friction coefficient x Refrigerant circulation amount ÷ {2 x Refrigerant density x (Cross-sectional area of flow path)^2}
Refrigerant circulation amount = cooling (or heating capacity) capacity / evaporation latent heat ratio (or condensation latent heat ratio)
Therefore, the pressure loss is inversely proportional to the refrigerant density and the latent heat of vaporization ratio (or the latent heat of condensation ratio). Since the pressure loss is maximized when the refrigerant density is minimum, the pressure loss can be compared with the saturated gas density on the evaporation side.

FIG. 13 is a diagram for explaining the GWP total amount value of the first embodiment and the GWP total amount value of the comparative example. Comparative Examples 1, 2, and 3 shown in FIG. 13 are “direct expansion type” in which the indoor unit and the outdoor unit are connected by a refrigerant pipe, and heat is exchanged by expanding the refrigerant near the space to be air-conditioned. It is a comparative example in the case of "direct expansion type) cycle". In each comparative example, the first refrigerant side 1L and the second refrigerant side 1L of the intermediate heat exchanger 22 are removed from the refrigerant-filled volume 10L of Embodiment 1 of the present invention, and the refrigerant-filled volume 8L is replaced with only the first refrigerant. assumes an enclosed volume.

Comparative Example 2 is a case of a direct expansion cycle in which R32 is charged as the primary refrigerant species, and the GWP total amount value is estimated to be 675×0.45×8≈2433. Using this as a reference value, the specific GWP total amount value is defined as 100%.

In Comparative Example 1, in the case of a direct expansion cycle in which R410A is charged as the primary refrigerant species, the GWP total amount value is 8234, and the specific GWP total amount value is 338.4%.

In Comparative Example 3, in the case of a direct expansion cycle in which R290 is charged as the primary refrigerant species, the GWP total amount value is 8, and the specific GWP total amount value is 0.3%.

On the other hand, in Embodiment 1, R290 is used as the first refrigerant and R466A is used as the second refrigerant to form an indirect cycle. Ignoring the influence of the refrigerant density distribution, in the first embodiment, the GWP total amount value=4×0.24×5+733×0.58×5=2127. Compared with the total GWP value of the direct expansion cycle using the current R410A refrigerant of Comparative Example 1, in the first embodiment, the total GWP value can be reduced by approximately 74%.

Furthermore, when compared with the total GWP value of the R32 direct expansion cycle of Comparative Example 2, the total GWP value can be reduced by about 13%. Therefore, the combination of refrigerant types in Embodiment 1 can reduce the total GWP compared to using R32 indoors, and is more advantageous than Comparative Example 2, which imposes restrictions on the amount of refrigerant and an obligation to install a safety device.

On the other hand, the pressure loss generated in the pipe 11G of the second refrigerant circuit 3 has a greater influence than the pipe 11L. From the ph diagram during cooling operation in FIG. 2, a pressure difference between points D2 and E2 occurs as pressure loss in the second refrigerant circuit side cycle, and the evaporation temperature (ET ) and compressor suction pressure (A1 point pressure) decrease, compression input increases, and primary side cycle efficiency decreases. Since the effect of pressure loss occurring in the pipe 11G also occurs in the case of a direct expansion cycle, the pipe diameter should be widened so that the pressure loss is the same. It does not cause performance degradation due to the formula cycle.

Although the R466 refrigerant has a smaller latent heat of vaporization ratio than R32, it has a saturated gas density that is 30% higher.

On the other hand, the temperature difference between the first refrigerant and the second refrigerant generated in the intermediate heat exchanger 22 is a factor of performance deterioration due to the indirect cycle, and must be minimized. Therefore, in the present embodiment, both the cooling operation and the heating operation are designed to be counterflow heat exchange.

As described above, according to the air conditioner 1 using the refrigerant indirect cycle according to Embodiment 1, the total GWP value of the refrigerant sealed in the air conditioner can be reduced, thereby contributing to the prevention of global warming. In addition, since the loss that occurs is suppressed and each device can be efficiently operated, a high coefficient of performance COP can be maintained for an indirect refrigerant cycle air conditioner. In addition, the suction of gas refrigerant by the pump 23 is suppressed, further improving efficiency.

In the configuration of FIG. 1, the first tank 17 and the second tank 18 for degassing are provided on both sides of the pump 23 that can be reversely rotated. The tank 17 may not be provided. In the case of a device dedicated to heating operation, the second tank 18 corresponds to the "first tank".

Embodiment 2.
Embodiment 2 will explain a case where a four-way valve is used as the cooling/heating switching means of the second refrigerant circuit. 14 is a refrigerant circuit diagram during cooling operation of the air conditioner according to Embodiment 2. FIG. 15 is a refrigerant circuit diagram during heating operation of the air conditioner according to Embodiment 2. FIG.

As shown in FIGS. 14 and 15 , an air conditioner 201 according to Embodiment 2 includes a first refrigerant circuit 2 and a second refrigerant circuit 203 . The first refrigerant circuit 2 has the same configuration as in the first embodiment. Second refrigerant circuit 203 includes pump 223 instead of pump 23 in the configuration of second refrigerant circuit 3 shown in Embodiment 1, and further includes four-way valve 224 . The pump 223 only needs to rotate in one direction, and does not necessarily need to be configured to be able to rotate in the opposite direction.

In this case, the intermediate heat exchanger 22, the pump 23, the first tank 17, and the four-way valve 223 preferably form an intermediate heat exchange unit 220, which is normally arranged on the outdoor side.

In the second embodiment, the pump 223 rotates in the forward rotation direction during the cooling operation and during the heating operation, so that the gas venting tank may be arranged only in the first tank 17 on the suction side of the pump. The shown second tank 18 is not required.

As described above, since the four-way valve is used, the air conditioner 201 according to the second embodiment has the same effects as the air conditioner 1 according to the first embodiment, although it is disadvantageous in terms of cost compared to the first embodiment. be able to.

As described above, according to the air conditioner 201 according to Embodiment 2, the total GWP value of the refrigerant sealed in the air conditioner can be reduced, thereby contributing to the prevention of global warming. Also. Since each device can be efficiently operated while suppressing the loss that occurs, it is possible to maintain a high coefficient of performance COP for an indirect refrigerant cycle air conditioner. In addition, the suction of gas refrigerant by the pump 23 is suppressed, further improving efficiency.

Finally, referring to the drawings again, the first and second embodiments will be summarized.
Referring to FIG. 1, the present invention relates to an air conditioner 1 having an intermediate heat exchanger 22 configured to exchange heat between a first refrigerant and a second refrigerant. The air conditioner 1 includes a first refrigerant circuit 2 containing a first refrigerant and a second refrigerant circuit 3 containing a second refrigerant. The first refrigerant circuit 2 includes a compressor 10 that compresses a first gaseous refrigerant, an outdoor heat exchanger 13 that exchanges heat between the first refrigerant and the outside air, and an expansion circuit that decompresses and expands the first refrigerant. It includes a device 24 and a first flow path H1 through which the first refrigerant passes in the intermediate heat exchanger 22 . The second refrigerant circuit 3 includes a pump 23 that pressurizes and conveys the second refrigerant in a liquid state, a second flow path H2 through which the second refrigerant passes in the intermediate heat exchanger 22, and a flow path between the second refrigerant and the room air. Indoor heat exchangers 31 and 41 that exchange heat therebetween, and a first tank 17 for degassing provided in the middle of the first pipes 11L1 and 11L2 in the second refrigerant circuit 3 are included. The first pipes 11L1 and 11L2 are pipes connected to the pump 23 and through which the second refrigerant sucked by the pump 23 passes.

In this way, since the first tank 17 for gas venting is provided in the suction side pipe of the pump 23, the reduction in efficiency of the pump 23 due to gas being sucked into the pump 23 is suppressed, and the circulating amount of the second refrigerant is reduced. can be stabilized.

As shown in FIG. 3, preferably, the first tank 17 has a tank housing 17C having a cross-sectional area larger than the diameters of the first pipes 11L1 and 11L2. The first pipe is divided into a pipe 11L1 on the upstream side and a pipe 11L2 on the downstream side of the refrigerant flow during cooling operation. The end of the upstream pipe 11L1 in the first tank 17 is arranged at a position higher than the end of the downstream pipe 11L2.

Since the first tank 17 is configured in this manner, even if the gas refrigerant flows into the first tank 17 , it is likely to stay in the space 17B above the first tank 17 . Therefore, the second refrigerant flowing from the first tank 17 toward the pump 23 can be in a gas-free state.

Similarly, the second tank 18 has a tank housing 18C having a cross-sectional area larger than the diameters of the second pipes 11L3 and 11L4. The second pipe is divided into an upstream pipe 11L4 and a downstream pipe 11L3 for the refrigerant flow during heating operation. The end of the upstream pipe 11L4 in the second tank 18 is arranged at a position higher than the end of the downstream pipe 11L3.

The first refrigerant circuit 2 further includes a four-way valve 12 connected to the discharge port and the suction port of the compressor 10 . As shown in FIG. 1, during cooling operation, the four-way valve 12 allows the first refrigerant to flow from the discharge port of the compressor 10 to the outdoor heat exchanger 13 and the first refrigerant that has passed through the first flow path H1 to the compressor 10. configured to flow into the inlet of the As shown in FIG. 4, during heating operation, the four-way valve 12 allows the first refrigerant to flow from the discharge port of the compressor 10 to the first flow path H1, and the first refrigerant that has passed through the outdoor heat exchanger 13 to the compressor 10. configured to flow into the inlet of the As shown in FIG. 3, the first pipes 11L1 and 11L2 are pipes through which the second refrigerant sucked by the pump 23 passes during the cooling operation.

In this way, since the four-way valve 12 is provided in the first refrigerant circuit 2, the flow of the first refrigerant flowing through the first flow path H1 of the intermediate heat exchanger 22 is reversed during cooling operation and during heating operation. . During cooling operation, gas is more likely to be generated on the suction side of the pump 23 than on the suction side of the pump 23 during heating operation. A first tank 17 is arranged. Although two degassing tanks are arranged in FIG. 1, for example, when only one degassing tank is arranged, it is preferable to arrange it on the suction side of the pump 23 during cooling operation.

More preferably, pump 23 described in Embodiment 1 is configured to be able to reverse the direction of the second refrigerant flowing through second flow path H2 by switching the direction of rotation. The direction of rotation of the pump 23 is determined such that the first coolant flowing through the first flow path H1 and the second coolant flowing through the second flow path H2 flow in opposite directions.

In Embodiment 1, the direction of rotation of the pump 23 is determined so that the first refrigerant and the second refrigerant exchange heat in a countercurrent manner, so the degree of subcooling can be reliably ensured, and the refrigeration cycle is improved. Efficiency can be maintained.

The second refrigerant circuit 3 further includes a second tank 18 for degassing provided in the middle of the second pipes 11L3 and 11L4. The second pipes 11L3 and 11L4 are pipes connected to the pump 23 and through which the second refrigerant discharged by the pump 23 passes during cooling operation.

As shown in FIG. 3, the internal volume of the first tank 17 is larger than the internal volume of the second tank 18 .

During the heating operation, the dryness of the second refrigerant on the suction side of the pump 23 is less likely to increase than during the cooling operation. Therefore, the degassing effect during heating is an auxiliary function. Therefore, the internal volume of the second tank 18 may be smaller than the internal volume of the first tank 17 . As a result, the size of the second tank 18 can be reduced, and the entire apparatus can be made compact.

The pump 23 is configured to be capable of reversing the direction of the second refrigerant flowing through the second flow path H2 by switching the direction of rotation between cooling operation and heating operation. As shown in FIGS. 7 to 10, the pump 23 has a vortex impeller 23D in which a plurality of blades rotating around a drive shaft are arranged in the circumferential direction as an impeller. The impeller 23D has a blade shape with directionality so as to exhibit the maximum lift when rotated in the first rotation direction R1 indicated by the arrow in FIG. 7 with respect to the drive shaft. A first rotation direction R1 indicated by a solid-line arrow in FIG. 7 is a direction in which the impeller 23D rotates during heating operation.

As shown in FIG. 10, each blade 84 that rotates around the drive shaft of the impeller 23D has a first corner portion C1 on the front end side and a second corner portion C2 on the rear end side with respect to the first rotation direction R1. The radius of curvature of the second corner C2 is greater than the radius of curvature of the first corner C1.

With such a configuration, the lift when rotated in the first rotation direction R1 indicated by the solid line arrow in FIG. It will be larger than the lifting height when it is raised.

Preferably, the second refrigerant is a CF ₃ I single refrigerant or a mixed refrigerant containing CF ₃ I, as described in Embodiment 1 and its modifications.

In Embodiments 1 and 2, the refrigerant circuits as shown in FIGS. 1 and 3 have been described, but the first refrigerant circuit 2 on the outdoor side and the second refrigerant circuit 3 on the indoor side form closed circuits, respectively. , and the intermediate heat exchanger 22 may be configured to perform cascade heat exchange.

Further, in Embodiments 1 and 2, the value of the volume of the refrigerant circuit has also been simplified. However, if the volume of the second refrigerant circuit 3 occupies about half of the whole, and the volume ratio of the refrigerant pipe connecting the intermediate heat exchanger 22 and the outdoor heat exchanger 13 is not negligible. By making the second refrigerant of the indirect cycle nonflammable and reducing the GWP, the effect of reducing the total GWP value can be obtained. Therefore, there are no restrictions as long as the application uses a vapor compression refrigeration cycle, and it is applicable to various capacity ranges and applications such as RAC, PAC, and building air conditioners.

In the first and second embodiments, the air conditioners shown in FIGS. 1 and 3 are configured with a plurality of indoor heat exchangers 31 and 41 and flow control valves 33 and 43 in order to adjust the temperature for each room. A multi-model equipped with the indoor temperature control units 30 and 40 has been described, but the same effect can be obtained not only in the multi-model but also in the case of a single model.

The embodiments disclosed this time should be considered as examples and not restrictive in all respects. The scope of the present invention is indicated by the scope of the claims rather than the description of the above-described embodiments, and is intended to include all modifications within the meaning and scope equivalent to the scope of the claims.

1,201 air conditioner, 2 first refrigerant circuit, 3,203 second refrigerant circuit, 7,11,11G,11L1,11L2,11L3,11L4,11L,72,73 piping, 10 compressor, 12,224 four-way valve , 13 outdoor heat exchanger, 17 first tank, 17A, 17B, 18A, 18B space, 17C, 18C tank housing, 18 second tank, 20,220 intermediate heat exchange unit, 22 intermediate heat exchanger, 23,223 Pump 23A Motor 23B Pump housing 23C Cylinder 23D Impeller 23E Flow path 24 Expansion device 30, 40 Indoor temperature control unit 31, 41 Indoor heat exchanger 33, 43 Flow control valve 61 , 63, 68, 69, 75, 76 temperature sensor, 84 vanes, 100 controller, 102 processor, 103 memory, C1 first corner, C2 second corner, H1 first flow path, H2 second flow path.

Claims (6)

An air conditioner comprising an intermediate heat exchanger configured to exchange heat between a first refrigerant and a second refrigerant,
a first refrigerant circuit enclosing the first refrigerant;
A second refrigerant circuit in which the second refrigerant is enclosed,
The first refrigerant circuit is
a compressor for compressing the first refrigerant in gaseous state;
an outdoor heat exchanger that exchanges heat between the first refrigerant and the outside air;
an expansion device that decompresses and expands the first refrigerant;
In the intermediate heat exchanger, including a first flow path through which the first refrigerant passes,
The second refrigerant circuit is
a pump that pressurizes and conveys the second refrigerant in a liquid state;
In the intermediate heat exchanger, a second flow path through which the second refrigerant passes;
an indoor heat exchanger that exchanges heat between the second refrigerant and indoor air;
In the second refrigerant circuit, including a first tank for degassing provided in the middle of the first pipe,
The first pipe is a pipe connected to the pump and through which the second refrigerant sucked by the pump passes during cooling operation ,
The second refrigerant circuit further includes a second tank for degassing provided in the middle of the second pipe,
The second pipe is a pipe connected to the pump and through which the second refrigerant discharged by the pump passes during the cooling operation,
The pump is configured to be capable of reversing the direction of the second refrigerant flowing through the second flow path by switching the direction of rotation between the cooling operation and the heating operation,
The air conditioner , wherein the internal volume of the first tank is larger than the internal volume of the second tank . The first tank has a tank housing having a cross-sectional area larger than the diameter of the first pipe,
The first pipe is divided into an upstream pipe and a downstream pipe,
2. The air conditioner according to claim 1, wherein an end of said upstream pipe within said first tank is arranged at a position higher than said end of said downstream pipe. The first refrigerant circuit is
further comprising a four-way valve connected to the discharge port and the suction port of the compressor;
The four-way valve is
During the cooling operation, the first refrigerant is caused to flow from the discharge port to the outdoor heat exchanger, and the first refrigerant that has passed through the first flow path is caused to flow to the suction port,
During the heating operation, the first refrigerant is caused to flow from the discharge port to the first flow path, and the first refrigerant that has passed through the outdoor heat exchanger is caused to flow to the suction port,
2. The air conditioner according to claim 1, wherein said first pipe is a pipe through which said second refrigerant sucked by said pump passes during said cooling operation. 4. The pump according to claim 3, wherein the direction of rotation of said pump is determined so that said first refrigerant flowing through said first flow path and said second refrigerant flowing through said second flow path flow counter-currently to each other. Air conditioner. The pump has a vortex impeller in which a plurality of blades rotating around a drive shaft are arranged in a circumferential direction as an impeller,
The vortex impeller has a shape in which the blade shape is directional so as to exhibit the maximum lift when rotated in the first rotation direction with respect to the drive shaft,
The air conditioner according to claim 1, wherein said first rotation direction is a direction in which said vortex impeller rotates during said heating operation. Each blade that rotates around the drive shaft of the vortex impeller has a first corner on the front end side and a second corner on the rear end side with respect to the first rotation direction,
The air conditioner according to claim 5 , wherein the radius of curvature of the second corner is larger than the radius of curvature of the first corner.

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