Patent Publication Number: US-2022235946-A1

Title: Air Conditioning Apparatus

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application is a U.S. national stage application of International Patent Application No. PCT/JP2019/029838 filed on Jul. 30, 2019, the disclosure of which is incorporated herein by reference. 
    
    
     TECHNICAL FIELD 
     The present disclosure relates to an air conditioning apparatus. 
     BACKGROUND 
     There is a need for an air conditioning apparatus that uses various types of alternative refrigerants to HCFC (Hydro Chloro Fluoro Carbon) refrigerants believed to destroy the ozone layer, that operates with a coefficient of performance comparable to those of the HCFC refrigerants in an actual machine, and that is able to use the alternative refrigerants safely as operating refrigerants. 
     Japanese Patent Laying-Open No. 7-269964 (PTL 1) discloses an example of such an air conditioning apparatus. This air conditioning apparatus includes a flow path A enclosing a refrigerant “a” and a flow path B enclosing a refrigerant “b”, and includes an intermediate heat exchanger in which these refrigerants “a” and “b” exchange heat with each other. 
     PATENT LITERATURE 
     PTL 1: Japanese Patent Laying-Open No. 7-269964 
     There has been a pressing need to prevent global warming in recent years, with the adoption of the Kigali Amendment to the Montreal Protocol, and the EU F-gas Regulation. In refrigeration and air conditioning fields, it has been required to make the shift to refrigerants with lower global warming potentials (“GWPs” hereinafter). HFC (Hydro Fluoro Carbon) refrigerants such as R410A, which have been mainly used in a vapor compression refrigeration cycle that is a heat source generating device of an air conditioning apparatus, have very high GWPs of about 2000. For this reason, HC (Hydrocarbon) refrigerants such as propane, HFO (Hydro Fluoro Olefin) refrigerants, and mixed refrigerants mainly composed of those refrigerants have been proposed as candidate alternative refrigerants to the HFC refrigerants. 
     Propane, being a highly flammable refrigerant, is subject to restrictions of international standards (such as the IEC) and national standards (such as the ISO, the High Pressure Gas Safety Act) when used indoors in terms of installation of safety measures and the amount of enclosed refrigerant. Therefore, an indirect air conditioning system has been developed, in which a primary circuit for generating a heat source is disposed outdoors, a secondary circuit is formed indoors with a heat transfer medium different from that of the primary circuit, and the heat transfer medium that has been subjected to heat exchange in an intermediate heat exchanger is transferred to an indoor heat exchanger. 
     Generally, the indirect air conditioning system uses water as the heat transfer medium on the secondary circuit side, and uses a pump as a drive device. When sensible heat of water is used, a liquid flow rate increases as compared to when latent heat of refrigerant is used, and a pump for transferring water and a liquid pipe increase in size, together with an increase in power. Therefore, it has been difficult to replace a refrigerant pipe of a direct air conditioning apparatus using existing refrigerants with a water pipe. In addition, water has been an unsuitable heat transfer medium in a facility where it is prohibited to bring in a water pipe in the first place, for example, a data center where many large-scale computers are installed. 
     An air conditioning apparatus using R32 as a refrigerant for a primary circuit has a high theoretical coefficient of performance (hereinafter referred to as theoretical COP) determined by a thermophysical property value, but has a GWP value of 675, which is significantly higher than what is defined by international regulations. In addition, R32, having a low flammable property, requires safety measures when used indoors. Accordingly, refrigerants with HFO-based refrigerants mixed therein have been developed, which have a theoretical COP equivalent to that of R32, and also have a low GWP property in order to improve the GWP and the flammable property. 
     For an indirect air conditioning system using such refrigerants, the configuration described in Japanese Patent Laying-Open No. 7-269964 can be employed. However, with regard to an indirect air conditioning apparatus using a highly flammable propane refrigerant, or a low flammable HFO-based refrigerant, or a mixed refrigerant mainly composed of these refrigerants on the primary circuit side, and using another refrigerant for a heat transfer medium of a secondary circuit, no detailed studies have been conducted on a method for selecting a refrigerant on the secondary circuit side for reducing a total GWP value, and on cooling-heating operation switching means for providing efficient operation. 
     SUMMARY 
     The present disclosure has been made to solve the problem described above, and has an object to provide an air conditioning apparatus capable of reducing a total GWP value without reducing efficiency by a combination of appropriate refrigerants. 
     The present disclosure relates to an air conditioning apparatus. The air conditioning apparatus includes a first refrigerant circuit enclosing a first refrigerant, and a second refrigerant circuit enclosing a second refrigerant. The first refrigerant circuit includes a compressor configured to compress the first refrigerant in a gaseous state, an outdoor heat exchanger configured to exchange heat between the first refrigerant and outside air, an expansion device configured to decompress and expand the first refrigerant, and a first flow path through which the first refrigerant passes in an intermediate heat exchanger configured to exchange heat between the first refrigerant and the second refrigerant. The second refrigerant circuit includes a pump configured to increase a pressure of the second refrigerant in a liquid state and transfer the second refrigerant, a second flow path through which the second refrigerant passes in the intermediate heat exchanger, and an indoor heat exchanger configured to exchange heat between the second refrigerant and air blown into a room. At least one of the first refrigerant and the second refrigerant has a global warming potential lower than that of R32, and the second refrigerant has a lower flammable limit concentration higher than that of the first refrigerant. 
     According to the air conditioning apparatus of the present disclosure, a total GWP value of enclosed refrigerant can be reduced, thereby contributing to preventing global warming. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a refrigerant circuit diagram during cooling operation of an air conditioning apparatus according to Embodiment 1. 
         FIG. 2  is a p-h diagram of a vapor compression refrigeration cycle in  FIG. 1 . 
         FIG. 3  is a refrigerant circuit diagram during heating operation of the air conditioning apparatus according to Embodiment 1. 
         FIG. 4  is a p-h diagram of a vapor compression refrigeration cycle in  FIG. 3 . 
         FIG. 5  is a diagram showing manners of combination of a forward direction and a backward direction of a direction of a first refrigerant by a four-way valve  12  in a first refrigerant circuit  2 , and a forward direction and a backward direction of a rotational direction of a pump  23  in a second refrigerant circuit  3 . 
         FIG. 6  is a flowchart illustrating control of switching a refrigerant circulation direction performed by a controller  100 . 
         FIG. 7  is a diagram showing relations between used refrigerants and total GWP values in Embodiment 1 and Comparative Examples 1 to 3. 
         FIG. 8  is a diagram showing comparison results of total GWP values (with respect to an R32 direct expansion cycle) according to Embodiment 1 and its Modifications 1 to 3. 
         FIG. 9  is a diagram showing comparison results of total GWP values according to Embodiment 2, with respect to the R32 direct expansion cycle. 
     
    
    
     DETAILED DESCRIPTION 
     In the following, embodiments of the present disclosure will be described in detail with reference to the drawings. While several embodiments are described below, it has been intended from the time of filing of the present application to appropriately combine configurations described in the respective embodiments. Note that the same or corresponding portions are designated by the same symbols in the drawings and will not be described repeatedly. 
     Embodiment 1 
       FIG. 1  is a refrigerant circuit diagram during cooling operation of an air conditioning apparatus according to Embodiment 1.  FIG. 2  is a p-h diagram of a vapor compression refrigeration cycle in  FIG. 1 .  FIG. 3  is a refrigerant circuit diagram during heating operation of the air conditioning apparatus according to Embodiment 1.  FIG. 4  is a p-h diagram of a vapor compression refrigeration cycle in  FIG. 3 . 
     As shown in  FIG. 1 , an air conditioning apparatus  1  includes a first refrigerant circuit  2 , a second refrigerant circuit  3 , and a controller  100 . First refrigerant circuit  2  corresponds to an “outdoor side cycle,” “heat source side cycle” or “primary circuit.” Second refrigerant circuit  3  corresponds to an “indoor side cycle,” “use side cycle” or “secondary circuit.” 
     First refrigerant circuit  2  includes, as main elements, a compressor  10 , a four-way valve  12 , an outdoor heat exchanger  13 , an expansion device  24 , a first flow path H 1  of an intermediate heat exchanger  22 , and a pipe  7 . 
     Second refrigerant circuit  3  includes, as main elements, a second flow path H 2  of intermediate heat exchanger  22 , a pump  23 , indoor temperature control units  30 ,  40 ,  50 , and pipes  11 L,  11 G. Indoor temperature control units  30 ,  40 ,  50  are connected in parallel with one another between pipe  11 L and pipe  11 G. 
     Indoor temperature control unit  30  includes an indoor heat exchanger  31 , a fan (not shown) for blowing indoor air to indoor heat exchanger  31 , and a flow rate control valve  33  for controlling a flow rate of a second refrigerant. Indoor heat exchanger  31  exchanges heat between the second refrigerant and the indoor air. 
     Indoor temperature control unit  40  includes an indoor heat exchanger  41 , a fan (not shown) for blowing indoor air to indoor heat exchanger  41 , and a flow rate control valve  43  for controlling a flow rate of the second refrigerant. Indoor heat exchanger  41  exchanges heat between the second refrigerant and the indoor air. 
     Indoor temperature control unit  50  includes an indoor heat exchanger  51 , a fan (not shown) for blowing indoor air to indoor heat exchanger  51 , and a flow rate control valve  53  for controlling a flow rate of the second refrigerant. Indoor heat exchanger  51  exchanges heat between the second refrigerant and the indoor air. 
     While an air conditioning apparatus including three indoor temperature control units is illustrated by way of example in the present embodiment, any number of indoor temperature control units may be provided. 
     First refrigerant circuit  2  encloses a first refrigerant and forms a closed circuit. The first refrigerant used in Embodiment 1 is, for example, an HFC refrigerant such as R32, an HC refrigerant such as propane (R290). Second refrigerant circuit  3  encloses the second refrigerant and forms a closed circuit. The second refrigerant used in Embodiment 1 is, for example, a CF 3 I single refrigerant, or a mixed refrigerant such as R466A including CF 3 I. 
     A feature of the present embodiment is that a combination of the first refrigerant and the second refrigerant is used, with at least one of the first refrigerant and the second refrigerant having a global warming potential lower than that of R32, and the second refrigerant having a lower flammable limit concentration higher than that of the first refrigerant. 
     First refrigerant circuit  2  is initially described. Compressor  10  sucks and compresses the low-temperature and low-pressure gaseous first refrigerant, and discharges the high-temperature and high-pressure gaseous first refrigerant. Four-way valve  12  is provided downstream of a discharge port of compressor  10 , and is configured to switch a flow path of the refrigerant discharged from compressor  10  between during the cooling operation and during the heating operation. During the cooling operation in which the refrigerant is circulated in a direction indicated by dashed line arrows in  FIG. 1 , four-way valve  12  forms a flow path from compressor  10  toward outdoor heat exchanger  13 . During the heating operation in which the refrigerant is circulated in a direction indicated by solid line arrows in  FIG. 3 , on the other hand, four-way valve  12  forms a flow path from compressor  10  toward intermediate heat exchanger  22 . 
     During the cooling operation, outdoor heat exchanger  13  functions as a capacitor for condensing and liquefying the high-temperature and high-pressure gaseous first refrigerant and dissipating heat to outside air. Conversely, during the heating operation, outdoor heat exchanger  13  functions as an evaporator for evaporating and gasifying the low-temperature and low-pressure liquid first refrigerant and taking heat from outside air. Expansion device  24  decompresses and expands the refrigerant passing therethrough into a low-temperature and low-pressure refrigerant. An electronic expansion valve can be used, for example, as expansion device  24 . 
     Intermediate heat exchanger  22  is configured such that heat exchange takes place between the first refrigerant circulating through first refrigerant circuit  2  and the second refrigerant circulating through second refrigerant circuit  3 . In intermediate heat exchanger  22 , there are formed a first refrigerant gap flow path and a second refrigerant gap flow path. The first refrigerant gap flow path is provided with a low-temperature-side inlet/outlet and a high-temperature-side inlet/outlet, and the second refrigerant gap flow path is provided with a low-temperature-side inlet/outlet and a high-temperature-side inlet/outlet. By appropriately disposing the low-temperature-side inlet/outlet and the high-temperature-side inlet/outlet of the second refrigerant gap flow path with respect to the high-temperature-side inlet/outlet and the low-temperature-side inlet/outlet of the first refrigerant gap flow path, the first refrigerant and the second refrigerant form a counter flow. 
     A plate heat exchanger, which is generally advantageous in cost and efficiency, is preferably used as intermediate heat exchanger  22 . In the plate heat exchanger, for example, a plurality of rectangular corrugated plates are stacked, with a plurality of rectangular gaps formed between them. The plurality of rectangular gaps alternately serve as the first refrigerant gap flow path and the second refrigerant gap flow path in the stacked direction. 
     In first refrigerant circuit  2 , there is provided four-way valve  12  on a pipe that discharges the first refrigerant in a gaseous state from compressor  10 . Four-way valve  12  switches a circulation direction of the first refrigerant so as to direct the first refrigerant from the discharge port of compressor  10  toward outdoor heat exchanger  13  during the cooling operation, and to direct the first refrigerant toward intermediate heat exchanger  22  during the heating operation. 
     In second refrigerant circuit  3 , on the other hand, pump  23  is configured to be able to switch its rotational direction between a forward direction and a backward direction. Pump  23  switches a circulation direction of the second refrigerant so as to direct the second refrigerant in a liquid state from pump  23  to indoor heat exchangers  31 ,  41 ,  51  during the cooling operation, and to direct the second refrigerant in a liquid state from pump  23  to second flow path H 2  of intermediate heat exchanger  22  during the heating operation. 
     In this manner, during the cooling operation, the second refrigerant circulating through second refrigerant circuit  3  is cooled by the first refrigerant circulating through first refrigerant circuit  2 . During the heating operation, on the other hand, the second refrigerant circulating through second refrigerant circuit  3  is heated by the first refrigerant circulating through first refrigerant circuit  2 . 
     Pipe  7  connects compressor  10 , four-way valve  12 , outdoor heat exchanger  13 , expansion device  24 , and first flow path H 1  of intermediate heat exchanger  22 , and allows the first refrigerant to circulate through these components. Intermediate heat exchanger  22  and pump  23  form an intermediate heat exchange unit  20 , and is usually disposed outdoors. 
     Pump  23  is suitable means for increasing a pressure of refrigerant in a liquid state and transferring the refrigerant, and is of a pump type capable of rotating in forward and backward directions. Such pump types include, for example, a Wesco type (also referred to as a vortex type, a regenerative type), a screw type, and the like. In air conditioning apparatus  1  of Embodiment 1, pump  23  on the second refrigerant circuit  3  side increases a pressure of the second refrigerant in equal volume while the second refrigerant remains in a liquid state, and hence only a small theoretical amount of enthalpy change is needed. Therefore, power required in pump  23  can be kept low. 
     Indoor heat exchangers  31 ,  41 ,  51  each exchange heat between the second refrigerant passing therethrough and indoor air. Indoor heat exchangers  31 ,  41 ,  51  each function as a so-called evaporator during the cooling operation, and each function as a gas cooler during the heating operation. 
     Pipes  11 L,  11 G connect intermediate heat exchanger  22 , pump  23 , and indoor heat exchangers  31 ,  41 ,  51 . By connecting these components with pipes  11 L,  11 G, second refrigerant circuit  3  through which the second refrigerant can circulate is formed. On pipe  11 L connected to respective inlet sides of indoor heat exchangers  31 ,  41 ,  51 , flow rate control valves  33 ,  43 ,  53  are provided. By adjusting the opening degrees of flow rate control valves  33 ,  43 ,  53 , the flow rates of the second refrigerant passing through indoor heat exchangers  31 ,  41 ,  51  installed in respective rooms from pump  23  can be adjusted. 
     In cooling conditions of  FIG. 1 , pipe  11 L from a point A2 near an outlet of the intermediate heat exchanger to the inlets of indoor heat exchangers  31 ,  41 ,  51  via pump  23  is a pipe for mainly passing liquid refrigerant. Pipe  11 G from the outlets of indoor heat exchangers  31 ,  41 ,  51  to an inlet E2 of the intermediate heat exchanger is a pipe for mainly passing gas refrigerant. 
     Indoor temperature control units  30 ,  40 ,  50  formed of indoor heat exchangers  31 ,  41 ,  51  and flow rate control valves  33 ,  43 ,  53  are disposed in respective rooms, and can control the temperatures of indoor air. 
     The switching of the refrigerant circulation direction performed in each of first refrigerant circuit  2  and second refrigerant circuit  3  is performed by controller  100 . 
     Controller  100  includes a processor  102  and a memory  103 . Memory  103  is configured to include, for example, a ROM (Read Only Memory), a RAM (Random Access Memory), and a flash memory. The flash memory stores an operating system, an application program, and various types of data. Processor  102  controls the overall operation of air conditioning apparatus  1 . Controller  100  shown in  FIG. 1  is implemented through execution, by processor  102 , of the operating system and the application program stored in memory  103 . For execution of the application program, the various types of data stored in memory  103  are referenced. 
     Controller  100  may further include a reception device if the controller receives a signal from a remote controller. When a plurality of indoor temperature control units are arranged, such a reception device is provided to correspond to each of the plurality of indoor units. 
     If the controller is implemented as a plurality of control units, the processor is included in each of the plurality of control units. In such a case, the plurality of processors cooperate with one another to perform the overall control of air conditioning apparatus  1 . 
     In accordance with outputs from a pressure sensor, a temperature sensor and the like, controller  100  controls compressor  10 , expansion device  24 , pump  23 , flow rate control valves  33 ,  43 ,  53 , and rotational speeds of not shown fans attached to heat exchangers  31 ,  41 ,  51 . 
     Controller  100  switches the circulation direction of the first refrigerant in first refrigerant circuit  2  by four-way valve  12  between during the refrigerant operation and during the heating operation. In conjunction therewith, controller  100  switches the rotational direction of pump  23  in second refrigerant circuit  3  such that the second refrigerant exchanges heat with the first refrigerant in a counter flow in intermediate heat exchanger  22 , and is subcooled at a suction port of pump  23 .  FIG. 5  is a diagram showing manners of combination of a forward direction and a backward direction of the direction of the first refrigerant by four-way valve  12  in first refrigerant circuit  2 , and a forward direction and a backward direction of the rotational direction of pump  23  in second refrigerant circuit  3 . 
     In  FIG. 5 , with regard to four-way valve  12 , a direction in which the first refrigerant discharged from compressor  10  flows toward outdoor heat exchanger  13  is indicated as the “forward direction,” and a direction in which the first refrigerant discharged from compressor  10  flows toward intermediate heat exchanger  22  is indicated as the “backward direction.” With regard to the rotational direction of pump  23 , a direction in which the refrigerant delivered from pump  23  flows toward indoor heat exchangers  31 ,  41 ,  51  is indicated as the “forward direction,” and a direction in which the refrigerant delivered from pump  23  flows toward intermediate heat exchanger  22  is indicated as the “backward direction.” 
     A combination of the “forward direction” and the “forward direction” is selected during the cooling operation, and a combination of the “backward direction” and the “backward direction” is selected during the heating operation, such that the first refrigerant and the second refrigerant flow as a counter flow in intermediate heat exchanger  22 . A combination of the “forward direction” as one of the directions and the “backward direction” as the other direction causes a parallel flow which is inefficient, and hence its use is prohibited. 
     The operation is controlled while required capability is ensured under temperature environment conditions during the heating operation and during the cooling operation, respectively, basically in a similar manner to that of general control. In the embodiments of the present disclosure, however, the operation is performed by noting the following points 1) to 3) in particular so as to provide highly efficient operation. 
     In the first refrigerant circuit through which the first refrigerant circulates, the operation is performed such that 1) a degree of subcooling (SC) of the first refrigerant at an evaporator outlet is ensured, and 2) a degree of superheat (SH) of the first refrigerant at the evaporator outlet is ensured. 
     In the second circuit through which the second refrigerant circulates, on the other hand, the operation is performed such that 3) the second refrigerant to be sucked into pump  23  is maintained in a liquid phase state. This is important in ensuring not only efficiency but also pump reliability. 
     (Cooling Operation Control) 
     When air conditioning apparatus  1  configured as shown in  FIGS. 1 and 2  is set in a cooling operation mode, controller  100  controls the rotational speed of compressor  10  based on a signal from a temperature sensor  63  disposed near the outlet of intermediate heat exchanger  22 , so as to maintain constant an outlet temperature at a point A1 near the outlet of intermediate heat exchanger  22 . 
     Compressor  10  also controls the rotational speed of pump  23  based on a signal from temperature sensor  63 , so as to maintain constant a degree of subcooling of the second refrigerant near the outlet of intermediate heat exchanger  22  shown at point A2. 
     Compressor  10  also controls the opening degrees of flow rate control valves  33 ,  43 ,  53  based on signals from temperature sensors  67 ,  68 ,  69  disposed near the outlets of indoor heat exchangers  31 ,  41 ,  51 , so as to maintain constant the degrees of superheat at locations indicated by points D21, D22, D23 near the outlets of indoor heat exchangers  31 ,  41 ,  51 . 
     That is, when the degrees of superheat at points D21, D22, D23 sensed by temperature sensors  67 ,  68 ,  69  are higher than “target degrees of superheat” set in advance as targets, the opening degrees of flow rate control valves  33 ,  43 ,  53  are increased, and when the degrees of superheat are lower, the opening degrees of flow rate control valves  33 ,  43 ,  53  are reduced. In this manner, the opening degrees of flow rate control valves  33 ,  43 ,  53  are controlled such that the degrees of superheat near the outlets of indoor heat exchangers  31 ,  41 ,  51  are maintained constant. 
     According to air conditioning apparatus  1  in the present embodiment, during the cooling operation, the rotational speed of compressor  10  is controlled such that the degree of subcooling near the outlet of intermediate heat exchanger  22  is maintained constant, and the opening degrees of flow rate control valves  33 ,  43 ,  53  are controlled such that the degrees of superheat near the outlets of indoor heat exchangers  31 ,  41 ,  51  are maintained constant. As a result, the rotational speed of pump  23  is controlled depending on the total load of second refrigerant circuit  3 , and air conditioning apparatus  1  can thereby be operated at an efficient and optimal operational point depending on the load of each of indoor heat exchangers  31 ,  41 ,  51 . 
     In intermediate heat exchanger  22 , the first refrigerant flows in through a point D1 near the inlet in a low-pressure liquid state, changes in phase while taking heat from the second refrigerant, and flows out to A1 near the outlet in a low-pressure gaseous state. The second refrigerant, on the other hand, flows in through point E2 near the inlet in a low-pressure gaseous state, changes in phase while losing heat to the first refrigerant, and flows out to point A2 near the outlet in a low-pressure liquid state, to be sucked into pump  23 . At this time, the first refrigerant and the second refrigerant exchange heat in the form of a counter flow, thereby attaining highly efficient heat exchange. In addition, the second refrigerant to be sucked into pump  23  can be maintained in a liquid phase state, and the efficiency of pump  23  can be maintained at high level. 
     Since each device can be operated efficiently as described above, a coefficient of performance COP corresponding to energy consumption efficiency of the entire air conditioning apparatus  1  can be maintained at high level. 
     (Heating Operation Control) 
     When air conditioning apparatus  1  is set in a heating operation mode, controller  100  controls the rotational speed of compressor  10  based on a signal from a temperature sensor  61  disposed at point D1 near the outlet of intermediate heat exchanger  22 , so as to maintain constant an outlet temperature at a position indicated by point D1 near the outlet of intermediate heat exchanger  22 . 
     In second refrigerant circuit  3 , the opening degrees of flow rate control valves  33 ,  43 ,  53  are controlled such that outlet temperatures at points C21, C22, C23 near the outlets of three indoor heat exchangers  31 ,  41 ,  51  are maintained at their respective target temperatures. Controller  100  controls the opening degrees of flow rate control valves  33 ,  43 ,  53  based on signals from temperature sensors  74 ,  75 ,  76  disposed near the outlets of indoor heat exchangers  31 ,  41 ,  51 . 
     With the control as described above, air conditioning apparatus  1  can be operated at an efficient and optimal operational point depending on the load of each of indoor heat exchangers  31 ,  41 ,  51 . 
     What is important here, so as not to reduce the efficiency, is to ensure the degree of subcooling at point D1 in the first refrigerant circuit, and to maintain the second refrigerant to be sucked into pump  23  in a liquid phase state. 
     In intermediate heat exchanger  22 , the first refrigerant flows in through point A1 near the inlet in a high-pressure gaseous state, changes in phase while losing heat to the second refrigerant, and flows out to point D1 near the outlet in a high-pressure liquid state. The second refrigerant, on the other hand, flows in through point A2 near the inlet in a high-pressure liquid state, changes in phase while taking heat from the first refrigerant, and flows out to point E2 near the outlet in a high-pressure gaseous state. At this time, the first refrigerant and the second refrigerant exchange heat in the form of a counter flow, thereby allowing the primary circuit to ensure the degree of subcooling at a point D2, and maintaining the high efficiency of the refrigeration cycle. 
     In the second refrigerant circuit, a temperature at point A2 near the suction port of pump  23  is measured by a temperature sensor  62 , and the amount of circulation of the second refrigerant is controlled by the rotational speed of pump  23  and the opening degrees of flow rate control valves  33 ,  43 ,  53 , such that the second refrigerant to be sucked into pump  23  is maintained in a liquid phase state. 
     With the control as described above, each device can be operated efficiently, thereby further improving the coefficient of performance COP corresponding to energy consumption efficiency of the entire air conditioning apparatus  1 . 
       FIG. 6  is a flowchart illustrating control of switching the refrigerant circulation direction performed by controller  100 . Referring to  FIGS. 1 and 6 , in step S 1 , controller  100  determines whether an operation mode that has been set by a remote controller or the like is cooling or heating. 
     When cooling is determined in step S 1 , in step S 2 , controller  100  sets four-way valve  12  such that the first refrigerant circulates in the forward direction. Further, in step S 3 , controller  100  causes pump  23  to rotate forward such that the second refrigerant circulates in the forward direction. With such control, the first refrigerant and the second refrigerant circulate as indicated by the dashed line arrows in  FIG. 1 , causing the first refrigerant flowing through first flow path H 1  and the second refrigerant flowing through second flow path H 2  in intermediate heat exchanger  22  to form a counter flow. 
     When heating is determined in step S 1 , on the other hand, in step S 4 , controller  100  sets four-way valve  12  such that the first refrigerant circulates in the backward direction. Further, in step S 5 , controller  100  causes pump  23  to rotate backward such that the second refrigerant circulates in the backward direction. With such control, the first refrigerant and the second refrigerant circulate as indicated by the solid line arrows in  FIG. 3 , causing the first refrigerant flowing through first flow path H 1  and the second refrigerant flowing through second flow path H 2  in intermediate heat exchanger  22  to form a counter flow. 
     In the example described above, as means for switching between the cooling operation and the heating operation, the circulation direction of the first refrigerant in first refrigerant circuit  2  is switched by four-way valve  12  disposed after the discharge of compressor  10 . In second refrigerant circuit  3 , on the other hand, the circulation direction of the second refrigerant is switched by switching the rotation direction of pump  23  between the forward and backward directions By performing these types of switching in conjunction with each other, the second refrigerant to be sucked into pump  23  can be maintained in a liquid state, while the heat exchange in the form of a counter flow is maintained between the first refrigerant and the second refrigerant flowing through intermediate heat exchanger  22 . 
     (Explanation of Refrigerant Types that Reduce Total GWP Value) 
     In a vapor compressor refrigeration cycle of an air conditioning apparatus, a conversion to lower GWP refrigerants has been promoted as a first step while HFC refrigerants having high energy efficiency are maintained. The conversion is, for example, from R410A with a GWP of 2090 to R32 with a GWP of 675. However, this is a far cry from a reduction target for the total GWP value set by international regulations. 
       FIG. 7  is a diagram showing relations between used refrigerants and total GWP values in Embodiment 1 and Comparative Examples 1 to 3. 
     Embodiment 1 describes an example where a highly flammable refrigerant (classification: A3) is used as the first refrigerant and a non-flammable refrigerant is used as the second refrigerant. When propane (R290) is used as the first refrigerant, the GWP is as low as 4. However, highly flammable refrigerants are restricted by laws and regulations when used indoors. 
     Next, as the second refrigerant which is a medium for transferring heat toward a room, non-flammable and nontoxic R466A (a mixed refrigerant of R32, R125, CF 3 I) was used in anticipation of a building easily affected by water. R466A has a GWP of 733, which is equivalent to that of R32, and which is approximately one-third of that of R410A. 
     For example, a volume of 10 L of enclosed refrigerant is distributed such that 5 L is enclosed in first refrigerant circuit  2  and 5 L is enclosed in second refrigerant circuit  3 . Of the 5 L in first refrigerant circuit  2 , 1 L is enclosed in intermediate heat exchanger  22 . Of the 5 L in second refrigerant circuit  3 , it is assumed that 1 L is enclosed in intermediate heat exchanger  22 , 1 L is enclosed in a combination of indoor temperature control units  30 ,  40 ,  50 , and 1 L is enclosed in remaining pipes  11 . 
     In the following, the total GWP values (=Σ (GWP×refrigerant volume×density)) by the respective enclosed refrigerants are compared. Here, the effect of refrigerant density is represented as “×Σρ.” The refrigerant density is distributed, in an evaporative heat exchanger and a condensing heat exchanger, varying from a saturated liquid state to a saturated gaseous state. Here, it was assumed that the first refrigerant had a condensing temperature (CT) of 45° C. and an evaporating temperature (ET) of 10° C., a saturated liquid state density and a saturated gaseous state density were calculated for each type of refrigerants from a physical property value database (such as the REFPROP of the NIST), and an averaged value pave was used. 
     The total GWP value of each type of refrigerants was calculated as the total GWP value=GWP×refrigerant volume×average density pave. In the present embodiment, a specific total GWP value was calculated with respect to a specific total GWP value of an R32 refrigerant direct expansion type, to demonstrate the reduction effect by the refrigerant type in the secondary circuit. 
     A pressure loss and the amount of refrigerant circulation during a disturbed flow, on the other hand, are expressed by the following equations: 
       Pressure loss=coefficient of friction×amount of refrigerant circulation÷{2×refrigerant density×(flow path cross-sectional area){circumflex over ( )}2}
 
       Amount of refrigerant circulation=refrigeration capacity (or heating capacity)÷evaporation latent heat ratio (or condensation latent heat ratio)
 
     Therefore, the pressure loss is inversely proportional to the refrigerant density and the evaporation latent heat ratio (or the condensation latent heat ratio). Since the pressure loss is at a maximum when the refrigerant density is at a minimum, the pressure losses can be compared by the values of evaporation-side saturated gas densities. 
     Comparative Example 1, Comparative Example 2, Comparative Example 3 in  FIG. 7  are comparative examples employing a “direct expansion cycle” in which an indoor unit and an outdoor unit are connected with a refrigerant pipe, and refrigerant is expanded to perform heat exchange near a space to be air conditioned. In each comparative example, it is assumed that in a volume of 8 L of enclosed refrigerant, which results from subtracting 1 L on the first refrigerant side and 1 L on the second refrigerant side of intermediate heat exchanger  22  from the volume of 10 L of enclosed refrigerant in Embodiment 1 of the present disclosure, only the first refrigerant is enclosed. 
     Comparative Example 2 employs the direct expansion cycle enclosing R32 as the primary refrigerant type, and it can be estimated that the total GWP value is 675×0.45×8≈2433. With this as a reference value, a specific total GWP value=100% is defined. 
     Comparative Example 1 employs the direct expansion cycle enclosing R410A as the primary refrigerant type, where the total GWP value is 8234, and the specific total GWP value is 338.4%. 
     Comparative Example 3 employs the direct expansion cycle enclosing R290 as the primary refrigerant type, where the total GWP value is 8, and the specific total GWP value is 0.3%. 
     In contrast, in Embodiment 1, an indirect cycle is formed by using R290 as the first refrigerant and using R466A as the second refrigerant. Ignoring the effect of the refrigerant density distribution, the total GWP value is 4×0.24×5+733×0.58×5=2127 in Embodiment 1. When compared to the total GWP value of the direct expansion cycle using the existing R410A refrigerant in Comparative Example 1, the total GWP value can be reduced by approximately 74% in Embodiment 1. 
     Further, when compared to the total GWP value of the R32 direct expansion cycle in Comparative Example 2, the total GWP value can be reduced by approximately 13%. Therefore, the combination of the refrigerant types in Embodiment 1 can reduce the total GWP value compared to when R32 is used indoors, and is more advantageous than Comparative Example 2 which is restricted in terms of the amount of refrigerant and which is required to install a safety device. 
     On the other hand, a pressure loss that occurs in pipe  11 G of second refrigerant circuit  3  has a greater effect than that in pipe  11 L. It is seen from the p-h diagram during the cooling operation in  FIG. 2  that a pressure difference between point D2 and point E2 occurs as a pressure loss in the second refrigerant circuit side cycle, and correspondingly, an evaporating temperature (ET) and a compressor suction pressure (point A1 pressure) in the first refrigerant circuit side cycle decrease, and a compression input increases, resulting in reduced efficiency of the primary side cycle. The effect of the pressure loss that occurs in pipe  11 G also occurs in the direct expansion cycle. Thus, a pipe diameter may be increased so that the pressure losses become equal. Usually, a pipe diameter increased twofold or less is practically allowed, and thus does not cause performance degradation by the indirect cycle. 
     The R466 refrigerant has an evaporation latent heat ratio lower than that of R32, but has a saturated gas density which is 30% higher than that of R32. Thus, the R466 refrigerant has a pressure loss well within an allowable range, and poses no practical problems as long as it is used in second refrigerant circuit  3 . 
     In contrast, a temperature difference between the first refrigerant and the second refrigerant that occurs in intermediate heat exchanger  22  causes performance degradation by the indirect cycle, and needs to be minimized. In the present embodiment, therefore, a design was adopted to perform heat exchange in the form of a counter flow both in the cooling operation and the heating operation. 
     As described above, according to air conditioning apparatus  1  using the refrigerant indirect cycle in Embodiment 1, the total GWP value of refrigerant enclosed in the air conditioning apparatus can be reduced, thereby contributing to preventing global warming. In addition, because the loss that occurs can be suppressed, and each device can be operated efficiently as described above, a high coefficient of performance COP can be maintained for an air conditioning apparatus of a refrigerant indirect cycle. 
       FIG. 8  is a diagram showing comparison results of total GWP values (with respect to the R32 direct expansion cycle) according to Embodiment 1 and its Modifications 1 to 3. 
     Modification 1 of Embodiment 1 
     In Modification 1 of Embodiment 1, it is assumed that 100% CF 3 I is used as the second refrigerant. A feature of the non-flammable (classification: A1) and nontoxic CF 3 I is that it has a very low GWP of 0.4 (taken from IPCC documents). The CF 3 I refrigerant, on the other hand, has an average density twice as high as that of R32, but has a saturated gas density which is 15% lower than that of R32, and also has a lower evaporation latent heat ratio, and therefore has a drawback of increasing the pressure loss on the gas pipe side when used in second refrigerant circuit  3 . 
     To make the pressure losses in the gas pipes equal, the inner diameter of pipe  11 G needs to be increased twofold, and the cross-sectional area of pipe  11 G needs to be increased fourfold. Assuming that the inner diameter of pipe  11 L is not changed because the pressure loss in pipe  11 L through which the liquid refrigerant flows does not change, the volume×density effect of pipe  11 G is increased 2.5 times, which corresponds to 2.5 L (an increase by 1.5 L). 
     When an indirect cycle is formed by using R290 as the first refrigerant and using CF 3 I as the second refrigerant, the total GWP value is 7. When compared to the R410A direct expansion cycle in Comparative Example 1, the total GWP value can be reduced by approximately 99.9% in Modification 1 of Embodiment 1. When compared to the R32 direct expansion cycle in Comparative Example 2, the total GWP value can be reduced by approximately 99.7%. 
     Modification 2 of Embodiment 1 
     In Modification 2 of Embodiment 1, it is assumed that R515B (a mixed refrigerant of R1234ze: 91%, R227ae: 9%) is used as the second refrigerant. The non-flammable (classification: A1) and nontoxic R515B has a GWP of 292, which is approximately one half of that of R32, and approximately 15% of that of R410A. The R515B refrigerant, on the other hand, has an average density which is 30% higher than that of R32, but has a saturated gas density which is 50% lower than that of R32, and therefore has a drawback of increasing the pressure loss on the gas pipe side when used in second refrigerant circuit  3 . 
     To make the pressure losses in the gas pipes equal, the inner diameter of pipe  11 G needs to be increased twofold, and the cross-sectional area of pipe  11 G needs to be increased fourfold. Assuming that the inner diameter of pipe  11 L is not changed because the pressure loss in pipe  11 L through which the liquid refrigerant flows does not change, the volume×density effect of pipe  11 G is increased 2.5 times, which corresponds to 2.5 L (an increase by 1.5 L). 
     When an indirect cycle is formed by using R290 as the first refrigerant and using R515B as the second refrigerant, the total GWP value is 1074. When compared to the R32 direct expansion cycle in Comparative Example 2, the total GWP value can be reduced by 56% in Modification 2 of Embodiment 1. 
     Modification 3 of Embodiment 1 
     In Modification 3 of Embodiment 1, it is assumed that a mildly flammable refrigerant (classification: A2L) is used as the first refrigerant, and a non-flammable (classification: A1) low GWP refrigerant is used, as in Embodiment 1, as the second refrigerant. In this case, R32 having high performance in theoretical COP, pressure loss, and specific latent heat capacity is selected as the first refrigerant, and non-flammable (classification: A1) and nontoxic CF 3 I is selected, as in Modification 1, as the second refrigerant. 
     When the inner diameter of pipe  11 G is increased twofold and the cross-sectional area of pipe  11 G is increased fourfold to make the pressure losses in the gas pipes equal, the secondary circuit volume increases by 1.5 L, which corresponds to 6.5 L. 
     When an indirect cycle is formed by using R32 as the first refrigerant and using CF 3 I as the second refrigerant, the total GWP value is 1523. With respect to the R32 direct expansion cycle in Comparative Example 2 as 100%, the total GWP value can be reduced by 37%. Although the effect of reducing the total GWP value is lower than when R290 is used as the first refrigerant, R32 provides superior performance (theoretical COP, pressure loss, and specific latent heat capacity). Since the effect of the pressure loss of the gas refrigerant of CF 3 I is low and at an allowable level, Modification 3 of Embodiment 1 is suitable for obtaining a high COP. 
     As described above, if the second refrigerant can be replaced and configured with a low GWP refrigerant that has been rendered non-flammable as in Modifications 1 to 3 of Embodiment 1, the total GWP value of refrigerant enclosed in the air conditioning apparatus can be reduced to thereby contribute to preventing global warming, even in the case of 100% CF 3 I, or even in the case of a mixed refrigerant including CF 3 I. In addition, although it may be required to take measures such as increasing the thickness of the pipe so as to reduce the pressure loss in pipe  11 G of second refrigerant circuit  3 , a design can be adopted to provide efficient operation of each device, and hence a high coefficient of performance COP can be maintained for an air conditioning apparatus of a refrigerant indirect cycle. 
     Embodiment 2 
     Embodiment 2 is similar to Embodiment 1 in the assumed indoor side and outdoor side cycles, the cooling-heating switching means, and the first refrigerant type, but is different in that a low flammable refrigerant (classification: A2L) is used instead of the non-flammable one as the second refrigerant. This is intended to attain a lower total GWP value than in Embodiment 1. 
       FIG. 9  is a diagram showing comparison results of total GWP values according to Embodiment 2, with respect to the R32 direct expansion cycle. 
     R1123, selected here as the second refrigerant, has a GWP of 4, which is equivalent to that of R290. R1123 has a specific evaporation latent heat which is one half of that of R32, has a saturated gas density which is twice that of R32, and also has a pressure loss in a main gas pipe which is equivalent to that of R32, thus posing no practical problems. In the following, the total GWP values by the respective enclosed refrigerants are determined and compared in a manner similar to that of Embodiment 1. 
     In Embodiment 2, when an indirect cycle is formed by using R290 as the first refrigerant and using R1123 as the second refrigerant, the total GWP value is 14. With respect to the R32 direct expansion cycle in Comparative Example 2 as 100%, the total GWP value can be reduced by 99.4%. 
     In the case of the R290 direct expansion cycle in Comparative Example 3, the total GWP value is 8, which is a very small value. However, since the amount of a highly flammable refrigerant (classification: A3) in a room is regulated when used in a large capacity air conditioning apparatus or cooling equipment, the R290 direct expansion cycle is not generally used. In such a case, by using the secondary refrigerant indirect cycle as in Embodiment 2, an air conditioning apparatus with a very low total GWP value equivalent to that of the R290 direct expansion cycle can be attained, although safety measures are required such as a gas leakage detector and a ventilation system in a room. 
     As described above, according to the air conditioning apparatus using the secondary refrigerant indirect cycle in Embodiment 2, the total GWP value of refrigerant enclosed in the air conditioning apparatus can be reduced, thereby contributing to preventing global warming. In addition, because the loss that occurs can be suppressed, and each device can be operated efficiently, a high coefficient of performance COP can be maintained for an air conditioning apparatus of a refrigerant indirect cycle. 
     Modification 1 of Embodiment 2 
     In Modification 1 of Embodiment 2, it is assumed that AMOLEA® 150Y4 (a mixed refrigerant of R1123: 60%, R32: 21.5%, R1234fy: 18.5%) is used as the second refrigerant. 150Y4 is a nontoxic and mildly flammable substance (classification: A2L) with a GWP of 146, and is restricted by laws and regulations when used indoors. 
     When an indirect cycle is formed by using R290 as the first refrigerant and using 150Y4 as the second refrigerant, the total GWP value is 392. With respect to the R32 direct expansion cycle in Comparative Example 2 as 100%, the total GWP value can be reduced by 84%. 
     As described above, according to air conditioning apparatus  1  in Modification 1 of Embodiment 2, the total GWP value of refrigerant enclosed in the air conditioning apparatus can be reduced, thereby contributing to preventing global warming. In addition, because the loss that occurs can be suppressed, and each device can be operated efficiently, a high coefficient of performance COP can be maintained for an air conditioning apparatus of a refrigerant indirect cycle. 
     Lastly, Embodiments 1 and 2 are summarized with reference to the drawings again. 
     An air conditioning apparatus  1  shown in  FIGS. 1 and 3  includes a first refrigerant circuit  2  enclosing a first refrigerant, and a second refrigerant circuit  3  enclosing a second refrigerant. First refrigerant circuit  2  includes a compressor  10  configured to compress the first refrigerant in a gaseous state, an outdoor heat exchanger  13  configured to exchange heat between the first refrigerant and outside air, an expansion device  24  configured to decompress and expand the first refrigerant, and a first flow path H 1  through which the first refrigerant passes in an intermediate heat exchanger  22  configured to exchange heat between the first refrigerant and the second refrigerant. Second refrigerant circuit  3  includes a pump  23  configured to increase a pressure of the second refrigerant in a liquid state and transfer the second refrigerant, a second flow path H 2  through which the second refrigerant passes in intermediate heat exchanger  22 , and indoor heat exchangers  31 ,  41 ,  51  configured to exchange heat between the second refrigerant and air blown into a room. As shown in  FIGS. 7 to 9 , at least one of the first refrigerant and the second refrigerant has a global warming potential lower than that of R32, and as was described in Embodiments 1 and 2, the second refrigerant has a lower flammable limit concentration higher than that of the first refrigerant. For example, a highly flammable refrigerant (classification: A3) can be used as the first refrigerant, and a non-flammable refrigerant can be used as the second refrigerant. Alternatively, for example, a mildly flammable refrigerant (classification: A2L) can be used as the first refrigerant, and a non-flammable (classification: A1) refrigerant can be used as the second refrigerant. 
     As a result, the total GWP value can be reduced, and in the event of refrigerant leakage into a room, flammability can be reduced as compared to when a highly flammable refrigerant is used entirely. 
     Preferably, in first flow path H 1 , a direction in which the first refrigerant flows through first refrigerant circuit  2  is reversed between during cooling operation ( FIG. 1 ) and during heating operation ( FIG. 3 ). In conjunction therewith, a direction in which the second refrigerant flows through second flow path H 2  is reversed between during the cooling operation ( FIG. 1 ) and during the heating operation ( FIG. 3 ), such that the first refrigerant flowing through first flow path H 1  and the second refrigerant flowing through second flow path H 2  form a counter flow. 
     As a result, heat exchange takes place always in the form of a counter flow in intermediate heat exchanger  22 . Thus, reduction in efficiency of the heat exchange in intermediate heat exchanger  22  can be prevented even when the switching is made between the cooling operation and the heating operation. 
     More preferably, first refrigerant circuit  2  further includes a four-way valve  12  connected to a discharge port and a suction port of the compressor. Four-way valve  12  is configured to, during the cooling operation, pass the first refrigerant from the discharge port of compressor  10  to outdoor heat exchanger  13 , and pass the first refrigerant that has passed through first flow path H 1  to the suction port of compressor  10 . Four-way valve  12  is configured to, during the heating operation, pass the first refrigerant from the discharge port of compressor  10  to first flow path H 1 , and pass the first refrigerant that has passed through outdoor heat exchanger  13  to the suction port of compressor  10 . Pump  23  is configured to be able to switch a rotational direction of a built-in motor, to thereby reverse the direction in which the second refrigerant flows through second flow path H 2 . Instead of causing pump  23  to rotate backward, a switching valve such as another four-way valve may be provided, and this switching valve may be connected to a suction port and a discharge port of pump  23  to reverse the circulation direction of the second refrigerant. 
     Preferably, second refrigerant circuit  3  further includes flow rate control valves  33 ,  43 ,  53  configured to control flow rates of the second refrigerant flowing through indoor heat exchangers  31 ,  41 ,  51 . Air conditioning apparatus  1  further includes a controller  100  configured to control first refrigerant circuit  2  and second refrigerant circuit  3 . Controller  100  is configured to, during cooling operation, control a rotational speed of compressor  10  so as to maintain constant a degree of subcooling of the second refrigerant that has passed through second flow path H 2 , control opening degrees of flow rate control valves  33 ,  43 ,  53  so as to maintain constant degrees of superheat of the second refrigerant that has passed through indoor heat exchangers  31 ,  41 ,  51 , and control, depending on a sum of loads of indoor heat exchangers  31 ,  41 ,  51 , a flow rate of the second refrigerant discharged from pump  23 . 
     Preferably, intermediate heat exchanger  22  is a plate heat exchanger configured to exchange heat between the first refrigerant and the second refrigerant. 
     Preferably, as was described in Embodiment 1 and its modifications, the second refrigerant is a CF 3 I single refrigerant, or a mixed refrigerant including CF 3 I. 
     Preferably, as was described in Embodiment 2 and its modification, the second refrigerant is an HFC1123 single refrigerant, or a mixed refrigerant including HFC1123, and the second refrigerant has a global warming potential lower than that of R32. In this case, the total GWP value can be further reduced. 
     Although the refrigerant circuit as shown in  FIGS. 1 and 3  has been described in Embodiments 1 and 2, any other configuration may be employed in which first refrigerant circuit  2  on the outdoor side and second refrigerant circuit  3  on the indoor side each form a closed circuit, and cascade heat exchange takes place in intermediate heat exchanger  22 . 
     The values of the volumes of the refrigerant circuits have also been described in simplified form in Embodiments 1 and 2. As long as the volume of second refrigerant circuit  3  occupies approximately one half of the total volume, and the volume makes up a non-negligible significant proportion of the refrigerant pipe coupling intermediate heat exchanger  22  and outdoor heat exchanger  13 , then the effect of reducing the total GWP value as a whole can be reduced by making the second refrigerant of the indirect cycle non-flammable and by reducing the GWP of the second refrigerant. Thus, any applications using a vapor compression refrigeration cycle can be employed without limitations, and many different capability areas and applications such as an RAC, a PAC, and a building air conditioner are applicable. 
     In Embodiments 1 and 2, the air conditioning apparatus shown in  FIGS. 1 and 3  has been described as a multiple type in which indoor temperature control units  30 ,  40 ,  50  formed of the plurality of indoor heat exchangers  31 ,  41 ,  51  and flow rate control valves  33 ,  43 ,  53  are provided so as to adjust the temperature for each room. However, a similar effect is also obtained in the case of a single type. 
     It should be understood that the embodiments disclosed herein are illustrative and non-restrictive in every respect. The scope of the present disclosure is defined by the terms of the claims, rather than the description of the embodiments above, and is intended to include any modifications within the meaning and scope equivalent to the terms of the claims.